The Japan Fluid Power System Society

Proceedings of

The 11th JFPS International Symposium on Fluid PowerHAKODATE 2020
Oct. 12-13, 2021

The 11th JFPS International Symposium on Fluid Power 2020

One-off sessions
Organized Session | Organized session
Opening Ceremony | Ceremony

Automobile

Opening Ceremony
Chair: Kazuhisa Ito, Shibaura Institute of Technology[OS1-01] A Study on Rotor Behavior of a Gerotor Pump
〇Ryo Kojima1, Takahiro Takeno2, Hideki Yanada1,[OP-01] Welcome Message〇Kazushi SANADA1 （1.Yokohama National

Hiroshi Yokoyama1 （1.Toyohashi University ofTechnology, 2.Toyohashi University of

University）

Technology(Currently at Terumo Corporation)）[OP-02] Opening Address〇Yutaka TANAKA1 （1.Hosei University）[OS1-02] Modeling of Compact Car Active Suspension
with Pneumatically Controlled Variable Damping

Invited Lecture | Invited Lectures

Force Oil Damper and Variable Rigidity Air

Invited Lecture 1

Spring〇Yasukazu Sato1, Shogo Sato1 （1.Yokohama[IL1-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute ofTechnology）

National University）[OS1-03] Power Absorber by Air Compression andRelease Cycle with Controller for Adjusting Air[IL-01] Piezoelectric pumps for hydraulic actuation〇Andrew Plummer1 （1.University of Bath）

Release Timing to Obtain Variable Absorbing
Capacity (First Report – Basic Concept and
Design)Invited Lecture | Invited Lectures〇Takashi Shibayama1, Toshinori Fujita1, Sho Ueyama1

Invited Lecture 2（1.Tokyo Denki University）[IL2-Introduction] Introduction[OS1-04] Clarification of Parameters and Development of〇Kazuhisa Ito1 （1.Shibaura Institute ofa Method for Estimating Loading Forces Acting

Technology）

on the Spool Valve of a Hydraulically Controlled[IL-02] Research on high performance electro-hydrostatic
actuator (EHA) system

Automotive Transmission〇Daisuke Yanagawa1, Masahiro Kouya1, Sho Tozuka1,

1〇Jiao Zongxia （1.Beihang University）

Masaru Shimada1, Idris Tengku1, Naoki Uezono1（1.Jatco,Ltd）

Invited Lecture | Invited Lectures

Invited Lecture 3

Organized Session | Organized session

Functional Fluid[IL3-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute ofTechnology）[OS2-01] A Study on a Mathematical Model for AC
Electroosmosis Micropump[IL-03] Fluid Power: from Motion Control to Powertrain
Innovation

Yo Makita1, 〇Kazuhiro Yoshida1, Sang In Eom1, Joonwan Kim1 （1.Tokyo Institute of Technology）

1〇Zongxuan Sun （1.University of Minnesota）[OS2-02] Multi-Layered Disk Type of Electro-Rheological
Braking Device for Small Mobile Robot

Invited Lecture | Invited Lectures〇Takanori Togawa1, Yuta Sato1, Yutaka Tanaka1,

Invited Lecture 4

Jinghui Peng2 （1.Hosei University, 2.Harbin Institute[IL4-Introduction] Introductionof Technology）
1〇Kazuhisa Ito （1.Shibaura Institute ofTechnology）[IL-04] Tribology Research on Fluid Power in Japan Review[OS2-03] Lightweight &Twin-driven MR Fluid Actuator for
Haptic Devices〇Tetsumasa Takano1, Asaka Ikeda2, Akinori

and State of the Art

Yamaguchi1, Isao Abe2, Takehito Kikuchi2〇Toshiharu Kazama1 （1.Muroran Institute of（1.Graduate School of Engineering, Oita University,

Technology）

2.Faculty of Science and Technology, Oita University）©The Japan Fluid Power System Society

The 11th JFPS International Symposium on Fluid Power 2020

[OS2-04] Investigation of Constant Velocity Motion with

Nonlinear Hydraulic Arms.

Physical Interaction System for Long- Term Stay

〇Teruo Kato1, Satoru Sakai1, Ryo Arai1 （1.University

in Microgravity Spaceof Shinshu）

1

1[OS3-06] On the Analysis of Energy Behaviors in Hydraulic〇Tetsuhito Fujita , Katsuki Machida , Yusuke
1

1

2

Cylinder Dynamics via Modeling of Experimental

Shimoda , Manabu Okui , Rie Nishihama , Taro
1

Nakamura （1.Chuo University, 2.Research andExcavators

Development Initiative of Chuo University）〇Ryo Arai1, Satoru Sakai1, Akihiro Tatsuoka2[OS2-05] Manufacturing and Evaluation of Micro（1.Shinshu University, 2.Mitsubishi Heavy

Electrohydrodynamic Pumps with Different

Industries）

Scales and Similar Dimensions
Organized Session | Organized session〇Kotaro Okada1, Masahito Nishikawara1,2, Shunichi

Aqua Drive 1

Naito1, Hideki Yanada1, Hiroshi Yokoyama1（1.Toyohashi University of Technology, 2.Worcester
Polytechnic Institute）[OS4-1-01] Consideration of Thermal Stability of the
Ultra-Precision Water-Lubricated Spindle[OS2-06] Study on a Low Pressure and Flowrate Driving ofMicro Leg Joints for Soft Robots
1〇Dmytro Fedorynenko1, Yohichi Nakao2（1.Tohoku University, 2.Kanagawa University）

1

1

Koji Michishita , Kazuhiro Yoshida , 〇Joon-wan Kim（1.Tokyo Institute of Technology）[OS4-1-02] Multidisciplinary Design Optimization of a
Tortuous Path Trim for a Labyrinth Control
Valve

Organized Session | Organized session〇Runlin Gan1, Xukang Li1, Song Liu1, Baoren Li1

Simulation and Modeling（1.Huazhong University of Science &Technology,[OS3-01] A Study on the Virtual Simulation Model of an
Excavator Equipped with a Tiltrotator Based on
Simscape

FESTO Pneumatic Technology Center）[OS4-1-03] Design of Magnetostrictive Power Generation
Device from Pulsating Pressure in Hydraulic1

1〇Seongwoong Choi , Kyungsin Kwak , Yongseok
1

1

1

Pipeline by Using Water Hydraulic Cylinder

Kim , Kyoungkwan Ahn , Soonyong Yang

〇Kaito Miyashita1, Shouichiro IIO1, Tsuyoki（1.University of Ulsan）

TAYAMA2, Ryuichi ONODERA2, Shyota ABE2

[OS3-02] Modeling, Simulation, and Control of Blade
Pitch to Improve the Performance of a
Hydrostatic Wind Turbine（1.Shinshu University, 2.Tohoku Steel Co., Ltd.）[OS4-1-04] Flow Characteristics of a Cavitating Jet
through a Small Rectangular Orifice with

1

1

〇Neil Christopher Garcia , Kim Stelson

Different Aspect Ratios（1.University of Minnesota）〇Hironori Takei1, Kohei Terakawa1, Shouichiro Iio1,[OS3-03] Internal Flow and Hysteresis Characteristic of

Kotaro Takamure2, Tomomi Uchiyama2, Futoshi

the Poppet Type Pressure Control Valve

Yoshida3 （1.Shinshu University, 2.Nagoya〇Seiei Masuda1, Fumio Shimizu2, Masaki Fuchiwaki2,

University, 3.KYB Corporation）

2

Kazuhiro Tanaka （1.Control Systems Engineering
Department, IHI Corporation, 2.Graduate School ofComputer Science and Systems Engineering, Kyushu
Institute of Technology）

Organized Session | Organized session

Aqua Drive 2[OS4-2-01] Switching Control of Latex Balloon Expansion[OS3-04] Flow Patterns and Hysteresis Characteristic of a
Poppet Valveby using Fluidic Switching Valve mediatedwith Coanda Effect〇Naoki Hirose1, Seiei Masuda2, Fumio Shimizu3,
33Masaki Fuchiwaki , Kazuhiro Tanaka （1.Voith IHI
Paper Technology, 2.IHI Corporation, 3.Kyushu
Institute of Technology）[OS3-05] Numerical Study on Identification Input for©The Japan Fluid Power System Society〇Keita Kaneko1, Kenjiro Takemura1 （1.Keio
University）[OS4-2-02] Comparison of Model-Free Adaptive
Displacement Control and Model Predictive
Displacement Control for Tap-Water-Driven

The 11th JFPS International Symposium on Fluid Power 2020

Muscle Considering Load Variation during

Shinohara1, So Shimooka2 （1.Okayama University

Experimentsof Science, 2.Okayama University）
1

1〇Satoshi Tsuruhara , Ryo Inada, Kazuhisa Ito[OS5-1-02] Development of Pneumatic Variable Linear（1.Shibaura Institute of Technology）

Stepping Actuator and Soft Stepping Actuator[OS4-2-03] Experimental study on Dual-Layer Type Vortex

with Bending Motion for Rehabilitation Device

Cup Driven by Aqua Drive System
1

of Hip Joint

1〇Wataru Kobayashi , Tetsuya Akagi , Shujiro〇Kota Oe1, Tetsuya Akagi1, Shujiro Dohta1, Takashi

Dohta1 （1.Okayama University of Science）

Shinohara1, Wataru Kobayashi1, So Shimooka2（1.Okayama University of Science, 2.Okayama

Organized Session | Organized session

University）

Soft Actuator 2[OS5-1-03] Development of Extension Type Flexible[OS5-2-01] Pneumatic Source Proposal for Improving

Pneumatic Actuator with Displacement Sensor

Portability and Responsiveness of Artificial

Using Ring-shaped Magnet and Hall Sensor for

Muscle via Dimethyl Ether Phase Change andTetrahedral-type Soft Mechanism

Combustion〇Kenshiro Takeuchi1, Takumi Kobayashi1, Tetsuya〇Ryuto Enjo1, Manabu Okui1, Taro Nakamura1

Akagi1, Shujiro Dohta1, Takashi Shinohara1, Wataru（1.Chuo university）

Kobayashi1, So Shimooka2 （1.Okayama University[OS5-2-02] Development of Six-Legged Mobile Robot
Using Tetrahedral Shaped Pneumatic Softof Science, 2.Okayama University）[OS5-1-04] Development of Highly Durable Straight Fiber
Type Pneumatic Artificial Muscle with a

Actuators
1

1

1〇Kenta Hase , Tetsuya Akagi , Shujiro Dohta ,Double Structural Air Chamber

Takashi Shinohara , Wataru Kobayashi , So

〇Naoki Saito1, Daisuke Furukawa1, Toshiyuki

Shimooka2 （1.Okayama University of Science,

Satoh1, Norihiko Saga2 （1.Akita Prefectural

2.Okayama University）

University, 2.Kwansei Gakuin University）

1

1[OS5-2-03] Development of the Transfer System for[OS5-1-05] Evaluation of Lifting Motion with Non-wearing

Bedridden Elderly and Disabled People using

Type Pneumatic Power Assist Device ~Pneumatic Actuators

Comparison of Active and Passive Type ~〇Feifei Cho , Keisei Kato , Ryota Endo , Takumi〇Masashi Yokota1, Reito Hirabayashi1, Masahiro

Kobayashi2, Tetsuya Akagi2 （1.National Institute of

Takaiwa2 （1.Tokushima University, Graduate

Technology Tsuyama College, 2.Okayama University

School Advanced Technology and Science,of Science）

2.Tokushima University, Graduate School of

1

1

1

[OS5-2-04] Development of Pneumatic Drive Pipe
Inspection Robot using Radial Bending TypeSoft Actuator〇Takashi Shinohara1, Hikaru Furuya1, Tetsuya
Akagi1, Shujiro Dohta1, Takumi Kobayashi1, So
2

Technology, Industrial and Social Sciences）
Genaral Session | Pneumatics

Robotics and Mechatronics 1[GS1-1-01] Development of a Simple Servo-Pneumatic

Shimooka （1.Okayama University of Science,

Three DOF Pick-And-Place Manipulator

2.Okayama University）

Chin-Yi Cheng1, Jyh Chyang Renn1, Shyang Jye
Chang1, 〇Ilham Saputra1 （1.National Yunlin

Organized Session | Organized session

Soft Actuator 1[OS5-1-01] Analysis and Design of Servo Valve Using

University of Science and Technology）[GS1-1-02] Development of Fingertip Mechanism With
Contact Point Estimation

Buckled Tubes for Desired Operation of〇Kei Mikami1, Kotaro Tadano1 （1.Tokyo InstituteFlexible Robot Arm

of Technology）〇Takumi Kobayashi1, Hideyuki Obayashi1, Tetsuya
Akagi1, Shujiro Dohta1, Wataru Kobayashi1, Takashi©The Japan Fluid Power System Society[GS1-1-03] Development of Outdoor Activity Assist Suit
〇Toshihiro Yoshimitsu1, Rui Matsumoto （1.

The 11th JFPS International Symposium on Fluid Power 2020
Sang-Ho Hyon1, 〇Tomoro Kai1 （1.Ritsumeikan

Kanagawa Institute of Technology）[GS1-1-04] Examination on Attitude Control System of

University）

Hand Manipulator with Compact Pneumatic[GS1-2-05] Comparison of Mechanical Drive System and

Cylinders by E-FRIT

Hydraulic Direct-Drive System for Motor

1

1

1

〇Shogo Tomita , Eiji Murayama , Yukio KawakamiPower（1.Shibaura Institute of Technology）〇Juri Shimizu1,2, Takuya Otani2, Kenji Hashimoto3,[GS1-1-05] Design and Fabrication of a Soft Filamentpolymer Jamming Actuator
1

Atsuo Takanishi2 （1.Hitachi, Ltd., 2.Waseda
University, 3.Meiji University）

1

1〇Peng Qin , Zhonghua Guo , MengYu Dou ,Genaral Session | Oil hydraulics

Zhongsheng Sun1, Yan Teng1, Xiaoning Li1

Energy Saving（1.Nanjing University of Science andTechnology(China)）[GS2-01] Research on the Characteristics of the Cylinder[GS1-1-06] Robotic Blood Vessel Mechanism for SelfHealing Function of Soft Robots〇Yuto Fujiwara1, Mitsuru Senoo1, Hiroyuki Asahara1

Kenjiro Tadakuma1, Shohei Inomata1, Yuta
2

（1.SMC Corporation）

2

2

Yamazaki , Fumiya Shiga , Masanori Kameoka , MD
2

1

Nahi Islam Shiblee , 〇Issei Onda , Tomoya
11[GS2-02] Design Guideline and Investigation of
Accumulator Parameters for a Novel Hybrid

1Takahashi , Yu Ozawa , Masahiro Watanabe ,2

Exhaust-Return Energy-Saving System

1

Architecture

Hidemitsu Furukawa , Masashi Konyo , Satoshi〇Seiji Hijikata1, Kazuhisa Ito2, Hubertus Murrenhoff1

Tadokoro1 （1.Tohoku University, 2.Yamagata（1.Institute for Fluid Power Drives and Systems (ifas),

University）

RWTH Aachen University, 2.Department of Machineryand Control Systems, College of Systems Engineering

Genaral Session | Oil hydraulics

and Science Shibaura Institute of Technology）

Robotics and Mechatronics 2[GS2-03] Research Regarding the Energy Saving[GS1-2-01] Design and Manipulability Analysis of a

Conditions of the Air Blow for Removing andRedundant Anthropomorphic HydraulicallyDrying Out Water

Actuated Manipulator〇Daisuke Kurakami1, Keiichirou Koga1, Gohei〇Fu Zhang1, Junhui Zhang1, Min Cheng2, Ruqi

Harimoto1 （1.SMC Corporation）

Ding3, Bing Xu1, Shen Zheng1 （1.State Key[GS2-04] Energy Regeneration and Reuse of Excavator

Laboratory of Fluid Power and Mechatronic

Boom System with Hydraulic Constantly Variable

Systems, Zhejiang University, 2.State Key Laboratory

Powertrainof Mechanical Transmissions, Chongqing University,〇Cuong Tri Do1, Kyoung Kwan Ahn1 （1.University of

3.Key Laboratory of Conveyance and Equipment,

Ulsan）

Ministry of Education, East China Jiaotong[GS2-05] Efficient Closed Pump Controlled Hydraulic-Gas
Balanced Energy Recovery Driving Method for

University）[GS1-2-02] Wide Field of View Projection Display for

Hydraulic Excavator Boom

Remote Control of Construction Robot〇Lianpeng Xia1, Long Quan1, Hongjuan Zhang1,〇Daisuke Kondo1 （1.Osaka University）

Yunxiao Hao1, Lei Ge1 （1.Taiyuan University of[GS1-2-03] Experimental Implementation of a Hydraulic
Turbine Access System with Six-DoF Active

Technology）[GS2-06] Experimental Validation of Improvement of the

Motion Compensation for Taiwan Offshore

Overall Efficiency for Electro-hydraulic Drive

Wind Farms

System using Efficiency Maps
1

1

Mao-Hsiung Chiang , Bo-Yen Chen , 〇Sheng-Chia〇Ha Tham Phan1, Yasukazu Sato1 （1.Yokohama

Lin1 （1.National Taiwan University）

National University）[GS1-2-04] Modular Hydraulic Servo Booster for MultiAxis Robots©The Japan Fluid Power System Society

Genaral Session | Pneumatics

The 11th JFPS International Symposium on Fluid Power 2020

Medical and Welfare[GS3-01] Development of Silicone Outer Shell Type
Pneumatic Soft Actuator〇Yuma Nakanishi1, Yasuhiro Hayakawa1, Keisuke[GS4-03] Feasibility and Precision Analysis of a Test Rig
with Adjustable Oil Film Thickness〇Haiji Wang1, Guanglin Shi1 （1.Shanghai Jiaotong
University）

Kida1, Hiroaki Ichii1 （1.National Institute of

Genaral Session | Pneumatics

Technology (Kosen) , Nara College.）

Components and Systems 1[GS3-02] Development of Pneumatically Driven
Verification System for Ophthalmic Needling
Operation
Feng Tao1, 〇Maina Sogabe2, Taro Ito3, Tetsuro
Miyazaki2, Toshihiro Kawase4,1, Takahiro Kanno5,
Yoshikazu Nakajima1, Kenji Kawashima2 （1.Tokyo
Medical and Dental University, 2.The University ofTokyo, 3.Tottori University, 4.Tokyo Institute ofTechnology, 5.RIVERFIELD Inc）

[GS3-03] Development of a Whole Body Training Device
by Multi-directional Force Input Using
Pneumatic Artificial Muscles
Soichiro Ito1, 〇Tetsuro Miyazaki2, Junya Aizawa3,
Toshihiro Kawase1,4, Maina Sogabe2, Takahiro Kanno5,
Yoshikazu Nakajima1, Kenji Kawashima2 （1.Tokyo
Medical and Dental University, 2.The University ofTokyo, 3.Juntendo University, 4.Tokyo Institute ofTechnology, 5.RIVERFIELD Inc.）[GS3-04] Development of Robotic Forceps Driven by Soft
Actuator with Built-In Displacement Sensor〇Osamu Azami1, Takahiro Kanno1, Toshihiro
Kawase2,3, Maina Sogabe4, Tetsuro Miyazaki4, Kenji
Kawashima4 （1.Riverfield Inc., 2.Tokyo Medical and
Dental University, 3.Tokyo Institute of Technology,
4.Tokyo University）[GS5-1-01] Magnetic Sensor Study for Improving Air
Turbine Spindle Performance〇Vanisara Kaewnamchai1, Tomonori Kato1, Kazuki
Kawakubo1, Kazumasa Yamashita1 （1.Fukuoka
Institute of Technology）[GS5-1-02] Study on Multi-Cylinder Type Wind Powered
Air Compressor
Applied Hypocycloid〇Ryota Tanoue1, Toshinori Fujita1 （1.Tokyo Denki
University）[GS5-1-03] Development of Bidirectional Arm Curl
Machine Using Pneumatic Artificial Rubber
Muscles〇Toshihiro Kawase1,2, Tomoya Nakanishi1, Shintaro
Yoshida3, Shingo Ohno3, Ryo Sakurai3, Tetsuro
Miyazaki4, Takahiro Kanno5, Maina Sogabe4,
Yoshikazu Nakajima1, Kenji Kawashima4 （1.Tokyo
Medical and Dental University, 2.Tokyo Institute ofTechnology, 3.Bridgestone Corporation, 4.The
University of Tokyo, 5.RIVERFIELD, Inc.）[GS5-1-04] Design of a Pneumatic Oscillator for Paper
Machine's Doctor Blade Systems
Stefano Colaiuda1, 〇Michele Gabrio Antonelli2,
Pierluigi Beomonte Zobel2, Massimiliano
Centofanti1 （1.SMC Italia S.p.A., 2.Department ofGenaral Session | Oil hydraulics

Industrial and Information Engineering andTribology, Seals, and Contamination Control

Economics, University of L'Aquila）[GS4-01] Effect of Sealing Surface Flatness on Leakage
Characteristics of Flange-Type Gasket Model

Genaral Session | Oil hydraulics

Components and Systems 2

Using Oil Viscosity-Temperature Relations〇Song Gao1, Toshiharu Kazama1 （1.MuroranInstitute of Technology）[GS4-02] Experimental Analysis of Rotational Motion of
Pistons and Slippers of a Swashplate Axial Piston
Pump Using Visualization Technique〇Takumi Furuya2, Toshiki Haga2, Toshiharu Kazama1（1.Muroran Institute of Technology, 2.Graduate
School of Muroran Institute of Technology）[GS5-2-01] Leakage Characteristics of a 3-port Pressure
Reducing Valve〇Haroon Ahmad Khan1, Byeong-Il Choi2, Jung-Ho
Park2, So-Nam Yun2 （1.University of Science andTechnology, Korea, 2.Korea Institute of Machineryand Materials）[GS5-2-02] Generation Mechanism of Flow Force Actingon Spool Valve〇Fumio Shimizu1, Kazuhiro Tanaka1 （1.Kyushu©The Japan Fluid Power System Society

The 11th JFPS International Symposium on Fluid Power 2020

Conventional Hydraulic Valve Controlled

Institute of Technology）[GS5-2-03] The Effect of the Spline Coupling on theArchitectures in Off-road Vehicles: The Case of a

Rotating Assembly Tilt Behavior in a High-Mini-Excavator

speed Axial Piston Pump

〇Mateus Bertolin1, Andrea Vacca1 （1.Purdue〇Haogong Xu1, Junhui Zhang1, Weidi Huang1, Bing
1

1

1

Xu , Fei Lyv , Xiaochen Huang （1.State Key

University）[GS6-04] Control of Air Bubble Content in Working Oil byLaboratory of Fluid Power and Mechatronic

Swirling Flow

Systems, Zhejiang University）

Sayako Sakama2, 〇Yutaka Tanaka1, Yasuhiro Kodera3,[GS5-2-04] First-order Trajectory Sensitivity Analysis ofMulti-level Pressure Switching Control System
1,2,3

Jing Yao

1

1, 〇Pei Wang , Xinhao Li , Yuwang

1

1

Yoshiaki Kitamura3 （1.Hosei University, 2.National
Institute of Advanced Industrial Science andTechnology, 3.KYB Corporation）

Cheng , Xiaoming Cao （1.School of Mechanical[GS6-05] Pulse Tests on Additive Manufactured ValveEngineering, Yanshan University, 2.Key Laboratory

Blocks – Damage Analysis and New Designof Advanced Forging &Stamping Technology andPossibilities

Science of Ministry of Education of China, 3.Hebei〇Sebastian Deuster1, Stefan Aengenheister1, Gunnar

Key Laboratory of Heavy Machinery Fluid Power

Matthiesen1, Katharina Schmitz1 （1.RWTH Aachen

Transmission and Control）

University, Institute for Fluid Power Drives and Systems[GS5-2-05] A Study on the Pulse Analysis and Vibration
Characteristics of Hydraulic System for
Prediction of Check Valve Behavior
1,2〇Jeong-Woo Park

1

, So-Nam Yun , Young-Bog

Ham1, Eun-A Jeong1 （1.Korea Institute ofMachinery &Materials, 2.Chungnam National
University）[GS5-2-06] Research on an Oil-hydraulic Component toReduce Pressure Pulsation〇Yasuo Sakurai1, Misaki Hashimoto2, Moritaka
Maehara3, Norikazu Hyodo4 （1.Ashikaga University,
2.MITSUBA Corporation, 3.SAWAFUJI ELECTRIC
CO., LTD., 4.TOKYO KEIKI INC.）
Genaral Session | Oil hydraulics

Construction, Components and Systems[GS6-01] Operating Information Presentation for
Hydraulic Construction Robot〇Hironao Yamada1, Takahiro Ikeda1, Satoshi Ueki1,
Kazuma Shinkai2, Katsutoshi Ootsubo3 （1.Gifu
University, 2.Sumitomo Heavy Industries, Ltd., 3.Kinjo
Gakuin University）[GS6-02] Frequency Response Analysis of Parallel Link
Mechanism using Oil-hydraulic Cylinders ofTunnel Boring Machines
Ryo Yamamoto1, 〇Kazushi Sanada1, Shigeaki
Ashikaga2 （1.Yokohama National University,
2.Komatsu Ltd.）[GS6-03] A Power-Split Hybrid Transmission to Drive©The Japan Fluid Power System Society(ifas)）[GS6-06] The Potential in Fluid Power Systems for a
Sustainable Future〇Katharina Schmitz1 （1.RWTH Aachen University,
Institute for Fluidpower Drives and Systems (ifas)）[GS6-07] New Intelligent Hydraulic Power Control System〇Takahiro Urai1, Ken Shindo1 （1.Bosch Rexroth
corporation）

The 11th JFPS International Symposium on Fluid Power 2020

Opening Ceremony | Ceremony

Opening Ceremony
Chair: Kazuhisa Ito, Shibaura Institute of Technology[OP-01] Welcome Message〇Kazushi SANADA1 （1.Yokohama National University）[OP-02] Opening Address〇Yutaka TANAKA1 （1.Hosei University）©The Japan Fluid Power System SocietyOP-01

The 11th JFPS International Symposium on Fluid Power 2020

Welcome Message
On behalf of the Japan Fluid Power System Society, it is my great honor to announce the 11th
JFPS International Symposium on Fluid Power HAKODATE 2020 as a major academic
international symposium. I present thanks all of you who are involved in the symposium.
Academic researchers of fluid power community have a long history of exchanging theirfriendship and academic information. I believe that we must do our best to take over the history
and expand fluid power community for the next generation. Fluid power researchers haveacademic interests in oil-hydraulics, pneumatics, water hydraulics, and functional fluids.
Application of fluid power have spread to broad fields. I believe this symposium is a key
conference which summarizes recent trends on fundamental and application of fluid power.
This message is devoted to success of the 11th JFPS International Symposium on Fluid PowerHAKODATE 2020.

Sincerely yours,
12th, October, 2021
Kazushi SANADA, Dr. Eng., Professor
President, the Japan Fluid Power System Society
Dean of College of Engineering Science
Yokohama National University, Japan

Copyright (C) JFPS All Rights Reserved.

- OP-01 -OP-02

The 11th JFPS International Symposium on Fluid Power 2020

Opening Address
On behalf of Japan Fluid Power System Society (JFPS) and JFPS-HAKODATE2020 committee, I am pleased towelcome you to the 11th International Symposium on Fluid Power Hakodate2020 at Hakodate Arena. We have atrust that you will have a rewarding time at this conference. Hope you and your loved ones are doing fine evenunder this difficult situation regarding Covid-19 pandemic. Unfortunately, considering the current world situation,
we have decided to deliver the symposium basically online with pre-recorded video presentations.
The first this symposium was held at Tokyo Institute of Technology in 1989. Since then, the JFPS International
Symposium on Fluid Power has been held every three years. Originally, it was scheduled to be held in October
2020 last year, but it was postponed for one year due to the Covid-19 pandemic. The symposium will address
recent developments in fluid power technologies such as hydraulics, pneumatics, water hydraulics and functional
fluids, and present basic researches, applications and case studies.
We have received excellent full papers with very interesting topics on fluid power. After the strictly peer-review
process, there are 73 accepted papers for the symposium. I would like to thank the 33 reviewers for theircooperation in the symposium. All accepted papers will be presented in 12 sessions for pre-recorded presentations
on the website. From these 73 papers, the best paper awards, the best student paper awards, and the GFPS best
paper award will be selected and presented in the closing & awards ceremony.
Four invited papers have also been organized by the committee in a two-day period. I would like to thank Prof.
Andrew Plummer, University of Bath, Prof. Zongxia Jiao, Beihang University, Prof. Zongxuan Sun, Universityof Minnesota, and Prof. Toshiharu Kazama, Muroran Institute of Technology.
We have participants that represent academia, industry, and research institutions from Japan, China, Korea,
Germany, U.S.A., Taiwan, Italy and U.K. This conference indeed follows the established path of bringing together
engineers, scientists, and practitioners that are working on the fundamental aspects of fluid power. Through yourparticipation and interaction with other attendees, this will be a truly exciting, memorable and fruitful symposium.
Hakodate, where the symposium is being delivered this time, is Hokkaido’s third largest city and a traditional
harbor city. Hakodate has experienced notable influence from overseas. And Hokkaido is a famous place for the
Ainu culture. The Ainu are an indigenous people from the northern region of the Japanese archipelago. The Ainu
traditional pattern is designed in the header of first page of each paper. I would like to thank Mrs. Nobuko Tsuda
and her related people for their cooperation in using the Ainu pattern.
I appreciate to be supported by grant funding from, The Precise Measurement Technology Promotion Foundation,
NSK Foundation for the Advancement of Mechatronics, Japan Fluid Power Association, and SMC Corporation.
During several years, all the committee members cannot help but be pleased that they overcame many difficulties
and arrived at this opening of the symposium. I would like to thank a member of people who have made it possible
to put this conference together, Prof. Toshiharu Kazama and Mr. Hidehiko Shimamura of co-chairperson, Prof.
Yukio Kawakami of Steering chair, Prof. Kazuhiro Yoshida of Paper chair, Prof. Kazuhisa Ito of Program chair,Prof. Toshiharu Yoshimitsu of Exhibition chair, Prof. Kenjiro Takemura of Website chair, Prof. Kenji Kawashima
of Budget chair, Prof. Yasuo Sakurai of Secretary general, and all the other members of HAKODATE2020
Executive Committee.
I hope you find this symposium stimulating and get some time to get valuable information.
Sincerely yours,
From Hakodate,

Yutaka Tanaka
JFPS HAKODATE2020, General Chairperson
October 12, 2021.
.

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- OP-02 -The 11th JFPS International Symposium on Fluid Power 2020

Invited Lecture | Invited Lectures

Invited Lecture 1[IL1-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute of Technology）[IL-01]

Piezoelectric pumps for hydraulic actuation〇Andrew Plummer1 （1.University of Bath）©The Japan Fluid Power System SocietyIL-01

The 11th JFPS International Symposium on Fluid Power 2020

Piezoelectric Pumps for Hydraulic Actuation
Andrew PLUMMER**

University of BathUK

Professor Andrew Plummer received his Ph.D from the University of Bath in 1991, forresearch into control of electro-hydraulic systems. He worked as a research engineer for
Thales from 1990, developing flight simulator control technology, before joining the
University of Leeds in 1994. From 1999 until 2006 he was global control systems R&D
manager for Instron, manufacturers of materials and structural testing systems. He is
now Director of the Centre for Power Transmission and Motion Control, University ofBath, and has published 180 papers in the field of motion and force control, many
relating to electro-hydraulic servo-systems. Prof. Plummer is Past Chair of the Institutionof Mechanical Engineers Mechatronics Informatics and Control Group and also the UK
Automatic Control Council, and is Associate Editor of both the International Journal of
Fluid Power and Control Engineering Practice. He is Chair of the Global Fluid Power Society.

ABSTRACT
There is an increasing demand for small-scale hydraulic actuation systems, working at around the one hundred
watts range, for example in human-scale robot applications. Pump-controlled actuators, where motion iscontrolled by varying the flow delivered to an actuator by a dedicated pump, are potentially energy efficient andallow the well-known flexibility in system design and control provided by hydraulics. However, compact andeffective variable-displacement or variable-speed pumps for delivering in the region of 1 L/min and 100 bar are
not widely available. Piezoelectric pumps are a possible solution. A single-cylinder piezoelectric pump consistsof a piezoceramic actuator – either a stack or a bending actuator – vibrating a piston in a pumping chamberconnected to suction and discharge lines by valves which rectify flow. Piezopumps have been found to becommercially viable in some low power fluid movement applications (10mW order of magnitude), for example
in medicine. However, higher power piezopumps are still a matter of research. In this presentation I describe
developments in this field, including reviewing some previous and current research, and describing the technicalchallenges and possible solutions.

1

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- IL-01 -The 11th JFPS International Symposium on Fluid Power 2020

Invited Lecture | Invited Lectures

Invited Lecture 2[IL2-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute of Technology）[IL-02]

Research on high performance electro-hydrostatic actuator (EHA)system〇Jiao Zongxia1 （1.Beihang University）©The Japan Fluid Power System SocietyIL-02

The 11th JFPS International Symposium on Fluid Power 2020

Research on High Performance Electro-Hydrostatic Actuator(EHA) system
Zongxia JIAO**

Beihang University
China

Prof. Zongxia Jiao is from Beihang University, Beijing, China. He is currently directorof the Fluid Transmission and Control Branch of the China Mechanical Engineering
Society, honorary Director of the Electromechanical Branch of the China Aviation
Society, Changjiang distinguished professor, winner of the National Science Fund forDistinguished Young Scholars. He has been engaged in the theory, key technology andequipment research of electro-hydraulic servo systems and basic components. He has
also worked in the field of the reliability and high-performance control of the aircraft
hydraulic system, the key technology and equipment of high-efficiency braking system,high-reliability hydraulic system and electro-hydraulic servo control. Prof. Jiao has been
granted 2 National Technology Invention Awards (2nd class) and 1 National S&T
Progress Awards (2nd class) and has established the IEEE/CSAA Aviation Electromechanical Series
International Conference. He has published more than 300 papers including 6 ESI high cited papers.

ABSTRACT
More/All-electric aircraft are currently the development trend of aircraft. Electro-hydrostatic actuator (EHA) is ahighly integrated actuation in power-by-wire aircraft utility system, including an electrical motor, pump andcylinder. EHA has the advantages of easy maintenance and good controllability. But its dynamic behavior isrelatively low and the temperature increment is fast, which restrict the use of EHA. In order to overcome thesedisadvantages, a new design of EHA which features active load-sensitive has been developed. Active loadsensitive is achieved by using a special pressure control valve to actively control the variable pump. It adjusts
the displacement of pump according to load pressure feedback and position feedback intelligently, which keeps
the advantage of high efficiency but also overcomes the weakness of low rigidity. Especially in the high loadcondition, the temperature of the motor can be reduced significantly. Both simulation and experiment results
indicate the performance improvement of the EHA.

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- IL-02 -The 11th JFPS International Symposium on Fluid Power 2020

Invited Lecture | Invited Lectures

Invited Lecture 3[IL3-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute of Technology）[IL-03]

Fluid Power: from Motion Control to Powertrain Innovation〇Zongxuan Sun1 （1.University of Minnesota）©The Japan Fluid Power System SocietyIL-03

The 11th JFPS International Symposium on Fluid Power 2020

Fluid Power: from Motion Control to Powertrain Innovation
Zongxuan SUN**

University of Minnesota
USA

Zongxuan Sun is a Professor at the Department of Mechanical Engineering, Universityof Minnesota. He received his Ph.D. degree in Mechanical Engineering from University
of Illinois at Urbana-Champaign in 2000. He is currently Director of the Engineering
Research Center for Compact and Efficient Fluid Power (CCEFP). He was a Staff
Researcher (2006-2007) and a Senior Researcher (2000-2006) at General Motors
Research and Development Center in Warren, MI. His research interests include controls,
fluid power, and mechatronics with applications to automotive and commercial vehicle
propulsion systems. Dr. Sun has published over one hundred fifty referred technicalpapers and received twenty two US patents. Dr. Sun is a recipient of the Charles E.
Bowers Faculty Teaching Award, George W. Taylor Career Development Award from
College of Science and Engineering, University of Minnesota, NSF CAREER Award,
SAE Ralph R. Teetor Educational Award, Best Paper Award from ASME Automotive and Transportation
Systems Technical Committee in 2018, Best Paper Award from 2012 International Conference on Advanced
Vehicle Technologies and Integration, Inventor Milestone Award, Spark Plug Award, and Charles L. McCuen
Special Achievement Award from GM R&D.

ABSTRACT
Fluid power is widely used in off-road vehicles including construction machinery and agriculture equipment for
driving and working functions. Given the increasing global focus on energy efficiency and environmental
impact, powertrain innovation for off-road vehicles is needed. Inspired by this urgent need, we would like torefocus on the fundamental strengths of fluid power and connect them to the energy conversion, power transfer
and motion control. This talk will present the study of motion control using fluid power, including accuracy,efficiency, and control bandwidth. Based on the motion control, we will discuss the powertrain innovation using
fluid power. Examples will be provided including free piston engine, camless engine, controlled-trajectory rapid
compression and expansion machine.

1

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- IL-03 -The 11th JFPS International Symposium on Fluid Power 2020

Invited Lecture | Invited Lectures

Invited Lecture 4[IL4-Introduction] Introduction〇Kazuhisa Ito1 （1.Shibaura Institute of Technology）[IL-04]

Tribology Research on Fluid Power in Japan Review and State of
the Art〇Toshiharu Kazama1 （1.Muroran Institute of Technology）©The Japan Fluid Power System SocietyIL-04

The 11th JFPS International Symposium on Fluid Power 2020

Tribology Research on Fluid Power in Japan⎯Review
and State of the Art
Toshiharu KAZAMA**

Muroran Institute of Technology
Japan

Toshiharu Kazama is a Professor of the Muroran Institute of Technology since 2006. He
joined the Institute as an Associate Professor in the department of mechanicalengineering in 1994. Prior to coming to the Muroran Institute of Technology, he hasdone research work as a Research Associate at Yokohama National University, Japan.
Dr. Kazama received his BEng, MEng, and PhD degrees from Yokohama National
University. Primary areas of research have been tribology and cavitation as related to
fluid power systems and components as well as machine elements and design. He has
worked in the fields of internal flow of hydraulic pumps and motors, lubrication ofhydrodynamic and hydrostatic bearings, sealing of packing and gaskets, and cavitationof liquid jets and erosion.

ABSTRACT
Hydraulic pumps, motors, and cylinders are the major components in hydraulic systems, which achievetransformation between mechanical energy and fluid pressure. The reliability, efficiency, and performance ofsuch a system depend on the tribology at the interface between its sliding, bearing, and sealing parts. This talkbriefly describes the short history of research on equipment tribology in Japan and surveys the knowledge andinsights obtained in the last couple of decades, including the activities of research committees in Japan Fluid
Power Systems Society (JFPS), and Japan Hydraulic and Pneumatic Society (forerunner of JFPS). With theliterature review alongside the cutting-edge technology, theoretical and experimental approaches for thesesystems in Japan are surveyed. Additionally, our results and efforts mainly regarding pump tests and slipper
models in terms of thermal effects, including liquid physical properties, thermohydrodynamic lubrication, andelastic deformation, are presented. Results on roughness effects, including mixed lubrication and surface
interaction, are also presented. Finally, challenges in fluid power tribology, future development agenda, andvision, with remarks on scope and perspective, are discussed.

1

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- IL-04 -The 11th JFPS International Symposium on Fluid Power 2020

Organized Session | Organized session

Automobile[OS1-01] A Study on Rotor Behavior of a Gerotor Pump
〇Ryo Kojima1, Takahiro Takeno2, Hideki Yanada1, Hiroshi Yokoyama1 （1.Toyohashi Universityof Technology, 2.Toyohashi University of Technology(Currently at Terumo Corporation)）[OS1-02] Modeling of Compact Car Active Suspension with PneumaticallyControlled Variable Damping Force Oil Damper and Variable Rigidity Air
Spring〇Yasukazu Sato1, Shogo Sato1 （1.Yokohama National University）[OS1-03] Power Absorber by Air Compression and Release Cycle with Controller for
Adjusting Air Release Timing to Obtain Variable Absorbing Capacity (First
Report – Basic Concept and Design)〇Takashi Shibayama1, Toshinori Fujita1, Sho Ueyama1 （1.Tokyo Denki University）[OS1-04] Clarification of Parameters and Development of a Method for Estimating
Loading Forces Acting on the Spool Valve of a Hydraulically Controlled
Automotive Transmission〇Daisuke Yanagawa1, Masahiro Kouya1, Sho Tozuka1, Masaru Shimada1, Idris Tengku1, Naoki
Uezono1 （1.Jatco,Ltd）©The Japan Fluid Power System Society

OS1-01

The 11th JFPS International Symposium on Fluid Power 2020

A Study on Rotor Behavior of a Gerotor Pump
Ryo KOJIMA*, Takahiro TAKENO**, Hideki YANADA*, Hiroshi YOKOYAMA**

Department of Mechanical Engineering, Toyohashi University of Technology
1-1, Hibarigaoka, Tempaku-cho, Toyohashi, Aichi, 441-8580 Japan(E-mail: kojima.ryo.lw@tut.jp)**
Department of Mechanical Engineering, Toyohashi University of Technology(Currently at Terumo Corporation)

Abstract. In our previous study, rotor behavior of a gerotor pump used for automatic transmissions was measured
using gap sensors, but the mechanism of the rotor behavior is yet to be made clear. In this study, in order to reveal themechanism of the rotor behavior, we measure the pressure distributions on both sides of the inner rotor and numericallysimulate the pressure distributions in the side clearances of the rotors. It was revealed that the pressure distributions
on both sides of the rotors were different and that such differences in pressure distribution caused axial forces andmoments of force being exerted on the rotors. It is concluded that the differences in pressure distribution between twosides of the rotors cause the tilting behavior of the inner rotor and the approach of the outer rotor to the casing sidewall.
Keywords: Gerotor pump, Rotor behavior, Pressure distribution, CFD

INTRODUCTION
In recent years, against the background of global warming, the fuel economy of automobiles has been attracting
more attention. Reducing the power loss of automatic transmissions has become more essential to improve fueleconomy. The power loss of a hydraulic pump incorporated in an automatic transmission occupies a relatively
high proportion of the total power loss of the automatic transmission. Therefore, decreasing the power loss of thepump, especially that resulting from friction torque, is vital to improve fuel economy [1].
The friction torque characteristics of hydraulic pumps have previously been examined [2-11], and a negative
resistance regime, in which the friction torque increases with decreasing rotational speed, is typically observed
and it becomes wider under increased pressures and temperatures [4, 7]. It is also known that solid-to-solid contact
takes place in this regime. However, for internal gear pumps, it has not been determined where the solid-to-solid
contact begins to take place.
Pham et al. [12, 13] experimentally investigated the eccentricity and tilt of the outer gear of a crescent internalgear pump/motor and found that the outer gear was in contact with the casing due to its tilt and eccentricity.
Uchino et al. [14] experimentally investigated the rotor behavior of a gerotor pump using gap sensors undervarious operating conditions. They revealed that the inner rotor rotated near the middle of the cover wall andcasing sidewall, slightly tilted with respect to them, and behaved in a manner roughly similar to precession, as
shown in Figure 1, and that the outer rotor rotated in the vicinity of the casing sidewall and did not vary its attitudewith rotational angle. In addition, it was revealed that as the dimensionless parameter 𝜇𝑁⁄∆𝑃 (where 𝜇 [Pa ∙ s] isthe oil viscosity, 𝑁 [s ] is the rotational speed, and ∆𝑃 [Pa] is the pressure difference) decreased, the clearance
between the outer rotor and the casing sidewall decreased and the eccentricity ratio of the outer rotor increased.
However, the mechanism of the rotor behavior has not yet been made clear.
In this paper, in order to reveal the mechanism of the rotor behavior of the gerotor pump, the pressure distributions
on both sides of the inner rotor are measured. In addition, numerical simulations are conducted to investigate thepressure distributions in the clearances between the rotors and their surrounding walls under a uniform side

Figure 1. Variation of inner rotor attitude during a single rotation measured at 60 C, 500 min , 1 MPa (arrow in (a)indicates the rotation direction), at (a) 0°, (b) 72°, (c) 144°, (d) 216°, (e) 288°, and (f) 342°. [14]1

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The 11th JFPS International Symposium on Fluid Power 2020

clearance condition and to compare the measured results with simulated ones. Based on the results, a mechanismof the rotor behavior is discussed.

EXPERIMENT
Test Pump
A schematic of the test gerotor pump used for automatic transmissions is shown in Figure 2 and the specificationsof the pump are described in Table 1. The displacement of the pump was experimentally determined. The sideclearances of the rotors and the diametral clearance of the outer rotor under no pressure were obtained bymeasuring the widths of the rotors, the outer diameter of the outer rotor, and the depth and internal diameter ofthe cylindrical hole of the casing. In order to measure the pressure distributions on both sides of the inner rotor,
pressure transduces (JTEKT, PYS-3-10-M, pressure range = 0 − 10 MPa) were mounted at the same radial andangular positions at the interval of 60 degrees on both the cover and cover-with-ports. Pressures in the sideclearances were transmitted to the pressure transduces through 1 mm diameter holes. In Figure 2(b), the angle
𝜃 = 0° was defined as the angular position at which a line starting from the center of the outer rotor passes throughthe center of the inner rotor, and the pumping stroke thus changes from a discharge stroke to a suction stroke at
𝜃 = 0° . Therefore, oil is sucked in the angular range of 𝜃 = 0° − 180° and is discharged in the range of 𝜃 =180° − 360° (0° ).

(a) Test pump schematic.

(b) Positions of pressure measurement.

Figure 2. Test pump schematic and positions of pressure measurement.
Table 1. Specifications of the test pump.
9.94

Displacement [cm3]
Dimensions / number of teeth

Inner rotor

Outer rotor

Width [mm]19.47619.479Tip diameter [mm]

38.0

33.4

Root diameter [mm]

28.2

43.7

Outer diameter [mm]―50.649Total side clearance [m]42

39

Diametral clearance [m]―149

Number of teeth

7

8

Experimental Apparatus and Method
Figure 3 shows the experimental apparatus. The test pump was driven by an induction motor at different rotational
speeds (500 − 4000 min ), discharge pressures (1 − 3 MPa), and oil temperatures (20 − 100°C), and the pumpdischarge pressure was adjusted using a throttle valve mounted at the outlet of the pump. The oil temperature wasmeasured in a reservoir. The viscosity of the oil was 0.0436 Pa ∙ s at 20°C , 0.0106 Pa ∙ s at 60°C , and
0.0043 Pa ∙ s at 100°C. The discharge pressure, suction pressure, and pressures in the side clearances of the inner

2

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The 11th JFPS International Symposium on Fluid Power 2020

rotor, rotational speed, discharge flow rate, and shaft torque were measured and recorded using a data logger witha sampling period of 1 ms.

Figure 3. Experimental apparatus.

NUMERICAL ANALYSIS
A simplified model was developed to numerically analyze the pressure distributions in the clearances around therotors and is shown in Figure 4, in which yellow spaces stand for the flow passages and are divided into six(1~3 MPa)
IN(a) Space between the inlet andsuction ports.

(b) Side gaps of rotors.

(c) Gap on the outer periphery of theouter rotor.

OUT (0 Pa)(d) Space between the two rotors.

(e) Space between the discharge ports andoutlet.

OUT (0 Pa)(f) Gaps between the driving shaft
and plain bearings.

(g) Virtual flow path.
Figure 4. Analytical model.

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The 11th JFPS International Symposium on Fluid Power 2020

subspaces. Figure (a) shows the space between the inlet port and the suction ports of the pump, Figure (b) the
space between the sides of the two rotors and cover/casing, Figure (c) the space between the casing circular walland the outer periphery of the outer rotor, Figure (d) the space between the two rotors, Figure (e) the space between
the discharge ports of the pump and the outlet, and Figure (f) the space between the driving shaft and plain bearings
machined in the cover and cover-with-ports.
The purpose of this numerical simulation is to examine whether some difference in pressure distribution is
generated between the two sides of the rotors or not. Therefore, to simplify the simulation, the following
conditions (1) to (4) were set: (1) The rotors do not rotate and are stationary in the center between the cover andcasing sidewall. (2) The rotors are not eccentric and do not tilt. (3) The pump is operated as a hydraulic motor,
though the rotors do not rotate. Therefore, to generate a pressure difference between the inlet and outlet, (4) a flow
path is formed between the teeth of the inner and outer rotors situated near the transition zone from the highpressure side to the low pressure side, as shown in Figure 4(g), and pressure in the range of 1 to 3 MPa is applied
to the inlet (Figure 4(a)) while 0 Pa is applied to the outlet (Figures 4(e), (f)). The height of the virtual flow path
was adjusted so that the theoretical flow rate of the pump corresponding to each rotational speed passed throughat each inlet pressure. A CFD software, ANSYS CFX (Ver.18.0), was used for the numerical simulation.

RESULTS AND DISCUSSION
Figure 5 shows the pressures measured at the points 1 to 6 shown in Figure 2(b) under the conditions of 60℃,
3 MPa, 3000 min . The highest pressure was observed at point 5 for this condition but was observed at point 6
for some other conditions. The magnitudes of the mean value and fluctuation of the pressure at each point weredifferent at the casing side and cover side. It is considered that such differences result from the difference in theside gap and from the inclination of the inner rotor as shown in Figure 1.

1 rotation

Pd
P1
P2
P3
P4
P5
P6

3.0

Pressure [MPa]

2.5
2.0
1.5
1.0

3.5

1 rotation

Pd
P1'
P2'
P3'
P4'
P5'
P6'

3.0
2.5

Pressure [MPa]

3.5

2.0
1.5
1.0

0.5

0.5

0.0

0.0
-0.5

-0.5
0

20

40
Times [ms]

60

0

80

(a)Casing side20

40
Times [ms]6080(b) Cover side

Figure 5. Pressures on the inner rotor sides measured at 60℃, 3 MPa, 3000 min (𝑃 is the discharge pressure, 𝑃 (𝑃 ) to𝑃 (𝑃 ) are the pressures at the points 1 to 6 shown in Figure 2(b)).

Figure 6 provides an example of the pressure distribution on the casing side and cover side of the rotors obtainedby the numerical simulation. The pressure distributions on both sides were almost the same, but it can be seen that
there is some difference in the area around 𝜃 = 180° . The length of the flow passage from the branching point
downstream the inlet to the casing-side suction port is shorter than that from the branching point to the cover-side
suction port, and the former is about a half of the latter (see Figure 4). The pressures at both suction ports aremacroscopically the same because the suction ports on both sides are connected through the space between thetwo rotors. Therefore, the pressure drop must be the same in both flow passages. This means that the flow velocity
in the shorter passage is about twice that in the longer passage, which was demonstrated by the numericalsimulation. Such a difference in inflow velocity at both suction ports may cause differences in velocity as well aspressure in the two side gaps of the rotors. In addition, all the fluid that have passed through the side gaps of therotors point toward the outlet. Such an asymmetric flow passage structure downstream the discharge port also
may cause asymmetric pressure distributions in the two side gaps of the rotors.

4

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The 11th JFPS International Symposium on Fluid Power 2020(a) Casing side.

(b) Cover side.

Figure 6. Simulated pressure distribution on both sides of rotors at 60℃, 3 MPa, 3000 min .

Figure 7 compares the measured pressure distributions on both sides of the inner rotor with simulated ones; forthe latter, the pressures at the points symmetrical to the measurement points with respect to the line connecting
the angular positions of 𝜃 = 0° and 𝜃 = 180° are plotted. An error bar attached to each measured value indicatesthe range of the pressure fluctuation at each measurement point, as shown in Figure 5. It is shown that thesimulated pressures were larger than the measured ones but that the tendency of the pressure variation with angle
was almost the same in the experiment and simulation. It is considered that the difference in the magnitude of thepressure between the experiment and simulation results from the differences in the axial position and the attitudeof the rotor. The inner rotor in the numerical simulation model were situated at the center between the cover andcasing sidewall and did not incline, while that in the test pump were not at the middle and inclined. The inclinationof the inner rotor causes the variations of the gaps and may increase the leakage flow rate through the gap, which
was confirmed by comparing the leakage flow rate between the experiment and numerical simulation. It isconsidered that a larger leakage flow rate in the experiment than in the numerical simulation generated a larger
pressure drop in the side gaps and that the pressures in the side gaps were lower in the experiment than in the
numerical simulation.

1
0.5
0
0

Pressure Pcov [MPa]

1
0.5

1

0.5

2.5

2.5

1.5
1
0.5
0

1.5

0

Exp-1MPa
Cal-1MPa

0

1.5

2
1.5

Exp-3MPa
Cal-3MPa2060 120 180 240 300 360
0 60 120 180 240 300 360
Angle θ [°]Angle θ [°](a) Pressure distribution on casing-side surface.

2.5
2Pressure Pcas [MPa]

1.5

2

2.5
Exp-2MPa
Cal-2MPa

Exp-2MPa
Cal-2MPa1
0.5

Pressure Pcov [MPa]2Pressure Pcas [MPa]2.5Exp-1MPa
Cal-1MPa

Pressure Pcov [MPa]Pressure P cas [MPa]2.5060 120 180 240 300 360
0 60 120 180 240 300 360
Angle θ [°]Angle θ [°](b) Pressure distribution on cover-side surface.

0

60 120 180 240 300 360
Angle θ [°]Exp-3MPa
Cal-3MPa

2
1.5
1
0.5
0
0

60 120 180 240 300 360
Angle θ [°]Figure 7. Comparison of pressure distribution on both sides of the inner rotor between experiment and numerical simulation
at 60℃, 3000 min .

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The 11th JFPS International Symposium on Fluid Power 2020

25
20
15
10
5
0
-5
-10
-15
-20
-25
0.01

1MPa
2MPa
3MPa

Fout [N]Fin [N]

Figure 8 shows the simulated axial force acting on the inner rotor 𝐹 and that acting on the outer rotor 𝐹 whichwere calculated from the simulated pressure distributions. Positive values of the forces mean that the forces act inthe direction from the cover to the casing sidewall. The numerical simulations were conducted under different
viscosities 𝜇, pressure differences ∆𝑃, and flow rates (corresponding to rotational speeds 𝑁 ), and the calculated
axial forces were plotted against the dimensionless parameter 𝜇𝑁⁄∆𝑃. It can be seen that the force 𝐹 tends topoint to the cover but that the force 𝐹 points to the casing sidewall at almost all the values of 𝜇𝑁⁄∆𝑃. The
latter result corresponds to the previously measured results [14], which showed that the outer rotor rotated in thevicinity of the casing sidewall. The magnitude of 𝐹 was larger than that of 𝐹 , which means that the pressure
difference between the two sides of the inner rotor is larger than that of the outer rotor. In addition to the axial
forces, the differences in pressure distribution on both sides of the rotors generate moments of forces acting on
the rotors. The moment of force acting on the inner rotor makes it incline because there was a relatively large
diametral clearance (65 m) between the hole at the center of the inner rotor and the driving shaft. The inclined
inner rotor may contact the cover and/or the casing sidewall. The pressure difference between the two sides mayvary with rotational angle because the side gaps varied with rotational angle in practice. Therefore, the momentof force acting on the inner rotor varies with rotational angle, which causes the variation of the inclination angle
with rotational angle, as shown in Figure 1.

0.1
μN/ΔP [-]

1
×10-6

25
20
15
10
5
0
-5
-10
-15
-20
-25
0.01

(a) Axial force acting on the inner rotor.

1MPa
2MPa
3MPa

0.1
μN/ΔP [-]1×10-6(b) Axial force acting on the outer rotor.

Figure 8. Simulated axial forces acting on the inner and outer rotors (𝐹 , 𝐹
casing sidewall.).

> 0 indicates that the force acts toward the

CONCLUSION
In this study, pressure distributions on both sides of the rotors of a gerotor pump used for automatic transmissions
were measured and numerically simulated using a simplified model. The measured and simulated pressure
distributions agreed qualitatively. The simulations showed that the pressure distributions on both sides of therotors were different even though the rotors were situated at the middle of the cover and casing sidewall and did
not incline, and that an axial force acting on the outer rotor toward the casing sidewall was generated, which
agreed with the result of the previous study [14].

ACKNOWLEDGMENTS
This work was supported by the Project for Building Simulation Platforms to Accelerate Development of NextGeneration Vehicles (Transmission Research Association for Mobility Innovation; TRAMI) grant in 2020.

REFERENCES
1. NAKAMURA, T., Improvement of Fuel Efficiency of Passenger Cars by Taking Advantage of Tribology, Journal ofJapanese Society of Tribologists, 2016, 61-2, pp.65-70. (in Japanese)
2. WILSON, W., E., Rotary-Pump Theory, Transactions of the ASME, 1946, 68-4, pp.371-384.
3. WILSON, W., E., Performance Criteria for Positive-Displacement Pumps and Fluid Motors, Transactions of the ASME,
1949, 71-2, pp.115-120.
4. INAGUMA, Y., HIBI, A., Friction Torque Characteristics of a Vane Pump, Journal of the Japan Hydraulics & Pneumatics
Society, 1994, 25-7, pp.843-849. (in Japanese)

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5. KLUGER, M., A., FUSSNER, D., R., ROETHLER, B., A Performance Comparison of Various Automatic Transmission
Pumping Systems, 1996, SAE paper 960424.
6. MANCO, S., NERVEGNA, N., RUNDO, M., ARMENIO, G., PACHETTII, C., TRICHILO, R., Gerotor Lubricating Oil
Pump for IC Engines, 1998, SAE Paper 982689.
7. INAGUMA, Y., Friction Torque Characteristics of an Internal Gear Pump, Proceedings of the IMechE, Part C: Journal of
Mechanical Engineering Science, 2011, 225-6, pp.1523-1534.
8. INAGUMA, Y., Oil Temperature Influence on Friction Torque Characteristics in Hydraulic Pumps, Proceedings of the
IMechE, Part C: Journal of Mechanical Engineering Science, 2012, 226-9, pp.2267-2280.
9. INAGUMA, Y., Friction of Outer Rotor Affecting Friction Torque Characteristics in an Internal Gear Pump, Proceedingsof the IMechE, Part C: Journal of Mechanical Engineering Science, 2015, 229-16, pp.3013-3026.
10. MANNE, V. H. B., VACCA, A., MERRILL, K., A Numerical Method for Evaluating the Torque Efficiency of HydraulicOrbit Motors Considering Deformation Effects and Frictional Losses, Mechanical Systems and Signal Processing, 2021,
146, 107051.
11. RITURAJ, R., VACCA, A., RIGOSI, M., Modeling and Validation of Hydro-Mechanical Losses in Pressure CompensatedExternal Gear Machines, Mechanism and Machine Theory, 2021, 161, 104310.
12. PHAM, T. H., MÜLLER, L., WEBER, J., Dynamically Loaded the Ring Gear in the Internal Gear Motor/Pump: Mobilityof Solution, Journal of Mechanical Science and Technology, 2018, 32-7, pp.3023-3035.
13. PHAM, T. H., MÜLLER, L., WEBER, J., Theoretical and Experimental Analysis of the Effect of Misaligned Ring Gear
on Performance of Internal Gear Motors/Pumps, Journal of Mechanical Science and Technology, 2019, 33-9, pp.40494060.
14. UCHINO, T., TAKENO, T., KOJIMA, R., YOKOYAMA, H., YANADA, H., Behaviors of Gears in an Internal Gear
Pump for Automatic Transmission, The Proceedings on Autumn Conference of Japan Fluid Power System Society 2019,2019, pp.35-37. (in Japanese)7

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Modeling of Compact Car Active Suspension
with Pneumatically Controlled Variable Damping Force Oil Damper
and Variable Rigidity Air Spring
Yasukazu SATO* and Shogo SATO***

Department of Mechanical Engineering, Faculty of Engineering, Yokohama National University
79-5, Tokiwadai, Hodogaya, Yokohama, Kanagawa, 240-8501 Japan(E-mail: sato-yasukazu-zm@ynu.ac.jp)**
Former Graduate Student, Graduate School of Engineering Science, Yokohama National University
79-5, Tokiwadai, Hodogaya, Yokohama, Kanagawa, 240-8501 Japan

Abstract. Active suspension system with a pneumatic linear actuator consisting of airtight devices, an air spring, anda variable damping force damper, has been developed with the aim of controlling the posture of a compact car during
turning and acceleration / deceleration. In this active suspension, the air pressure can control both the rigidity of the
air spring and the damping force characteristics of the damper. In the variable damping force damper, the relation
between the air pressure and the damping force has non-linear characteristics, and it is important to model the variable
damping force mechanism at the design stage in order to estimate the performance of the active suspension and thevehicle behavior. In this paper, the model of the fluid resistance element in the variable damping force mechanism isconducted for expressing its non-linear characteristics. This model fits actual damper behavior well.
Keywords: Modeling, Suspension, Air spring, Damper, Automobile

INTRODUCTION
Automobile suspension is a mechanism that generates passive spring force and damping force by coil spring andoil damper, and the specifications of spring and damper determines the suspension performance. Some kinds ofactive suspension system have an air spring actuator controlled by air pressure or a damper actuator controlled by
the auxiliary oil hydraulic power source. However, it is necessary to install control valves to the active suspension
system, and there is a problem of difficulty to install them in a small vehicle.
For this issue, the authors proposed a compact and lightweight airtight circuit pneumatic system in which actuators,
tanks, compressors, etc. were composed of airtight equipment that has no sliding parts and no leak of gas, andapplied it to active suspension using air springs [1]. Figure 1 is a photograph of a small vehicle equipped with an
air spring active suspension to control the attitude of roll behavior during turning and pitch behavior during
acceleration / deceleration. In this active suspension, the pressure-resistant thin-wall metal bellows developed by
the authors is used for the airtight variable volume actuator and compressor. Figure 2 shows the structure of thepressure-resistant thin-wall metal bellows. A viscoelastic ring is inserted in the inner diameter groove of aconventional thin-wall metal bellows to maintain the curvature of the groove against deformation of the bellows
and prevent stress concentration and buckling of the groove due to the rise in gas pressure inside the bellows.
Therefore, it maintains elasticity of thin-wall metal bellows, and its pressure resistance is 7 times that ofconventional thin-wall metal bellows. It is possible to output 7 times power in the same cross-sectional area by
supplying higher air pressure. The authors installed a pressure-resistant thin-wall metal bellows air spring on theouter cylinder of the damper to control the spring rigidity instead of the coil spring of the normal suspension, andFigure 1. Vehicle posture control using metal bellows air-spring active suspension

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also installed a metal bellows of the same structure inside the damper to vary the damping force by air pressure
control. As shown Fig.3, an active suspension with pneumatically controlled variable damping force damper has
been developed.
Since this variable damping force damper has a non-linear damping force characteristic with respect to thecontrolled air pressure, for applying it to various vehicles with different specifications, a model of the non-linear
damping force characteristic needs to be constructed. is constructed. At the design stage, it is important to estimate
the performance of the active suspension and the behavior of the vehicle. In this paper, the authors construct amodel of the fluid resistance element of the damping force device, and report that the characteristic representationby the model shows good agreement with the experimentally measured result of the prototype damper. Except forthe damping force device, the authors have reported the model of metal bellows air spring and conventional oil
damper in the former researches [1],[2].

Figure2. Pressure-resistant thin-wall metal bellows

Figure 3. Developed metal bellows air spring with
controllable damping force device

Figure 4. Combined piston with damping force devices

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STRUCTURE AND WORKING PRINCIPLE OF VARIABLE DAMPING FORCE
CHARACTERISTICS DAMPER
In the variable damping characteristic damper shown in Figure 3, the controlled air pressure is supplied to thethin-wall metal bellows chamber corresponding to the gas chamber of a normal oil damper,. The metal bellows
separates the oil inside the damper from the air in the bellows chamber.
With this structure, the controlled air pressure in the bellows chamber can adjust the pressure of the oil inside thedamper. The damper piston is a composite piston in which damping force generation mechanisms are installed on
the head side and rod side of the piston, respectively. The mechanism installed on the head side generates damping
force during contraction, and the mechanism on the rod side generates a damping force during extension. Figure
4 shows the damping force generators. Each contraction and extension damping force generator is symmetricallyarranged around the combined piston. The neutral pressure adjusted to the bellows chamber pressure acts as astatic pressure on the entire mechanism. The neutral pressure controls the set load of the disc spring by air volume
change of the backup air chamber kept airtight with the metal bellows. The set load of the disc spring is applied
to press the disc valve. As the neutral pressure rises, the set load of the disc spring increases. For contraction, theoil flow is squeezed by the disc valve (part C) through the gap (part A) and oil passage (part B) on the outer
periphery of the damping force generation mechanism to generate a pressure difference. The load is set to the discspring, and a corresponding damping force is generated. Then, the oil passes through the passage (part D, broken
line part in Figure 4) apart from the disc valve passage. The oil flows out from the rectifying valve (part E) witha low set load. Damping force is generated by the same manner of oil flow during extension.

MODEL CONSTRUCTION OF FLUID RESISTANCE ELEMENT OF DAMPINGFORCE GENERATOR
Principle of Actuation of Damping Force Generator
Figure 5 shows the measured values of the damping force characteristics. The air pressure of 0.2MPa, 0.4MPa,
and 0,6MPa controls the damping force characteristics. In the figure, the vertical axis is converted from the damper
piston speed to the flow rate Q passing though the piston, and the horizontal axis is converted from the damping
force to the pressure difference ΔP between both sides of the piston.
Based on the actual measurement results and the operating principle, the model of the damping force
characteristics of the variable damping force damper is constructed. In the condition that the pressing force on thedisc valve by the controlled air pressure is sufficiently large, the disc valve below the cracking pressure does not
lift. It is assumed that the thin parallel walls passage flow represents the clearance leakage flow region Q1 in thiscondition. The flow region Q2, in which the pressure difference at the piston ΔP increases and the disc valve
bends and lifts, is added to the region of Q1.

Figure 5. Measured damping force characteristics for damper rod extension（Damper force and velocity are converted to pressure and flow rates）

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Modeling of Fluid Resistance at Damping Force Generator
Leakage flow region – Disc valve in damping force generator does not lift
Both flow regions are expressed as the followings;
For the leakage flow region Q1;
𝑏ℎ
Δ𝑃
12𝜇𝑙

𝑄1

1

where the leakage flow rate Q1 is assumed as laminar flow, and symbols are defined as the clearance width b,clearance height h, the clearance overlap length l, and oil viscosity coefficient μ , respectively. In the flow region
Q1, the disc valve does not lift. Therefore, all parameters are fixed except the pressure difference P.
Orifice flow region – Disc valve in damping force generator lifts according to pressure difference
The formula of orifice is applied to the estimation of passing flow rate through the orifices consisting of thecircular hole and the disc valve in the damper piston. The orifice holes are circularly arranged in equal interval.
The orifice flow region Q2 is expressed by;
𝑄2

𝑛𝐶 𝐴

2∆𝑃
𝜌

2

where, symbols are defined as the number of orifice holes n, the discharge coefficient Cd, and oil density ρ,
respectively. The discharge coefficient Cd is represented by the function of Reynolds number, and the approximate
expression in the reference [3] is used. The cross-sectional area of flow path, A, for one orifice hole is assumed as
the cylindrical surface with the circumference of hole and the valve lift but bending of the disc valve from thecenter to the outer circumference is also considered. The thin wall disc valve lifts and bends, since the center ofthe disc is loaded with pneumatically controlled set force and the outer area of disc receives the pressure at thehole. Considering bending of disc valve, the cross-sectional area A is expressed by;
𝐴

𝜋𝑑𝛿 ∙

1
𝜋

3

where, symbols are defined as the diameter of the circular hole d, the lift of the disc valve at the outer peripheralside δ, respectively. In Eq.(3), 1 / π is the correction coefficient of the flow path area due to the disc valve bending.
The disc valve lift is expressed by;𝛿𝐴 ∆𝑃 𝐹𝑘4

where, A is the equivalent pressure receiving area of disc valve, which is considered not only circular hole but
also the pressure distribution around the hole. F0 is the initial load of the disc valve determined by the set pressureof air pressure control, and k is the combined spring constant of the disc valve and the disc spring. For the orificeflow region Q2, t the following equation is derived from Eq. (2) to Eq. (4);𝑄2𝑛𝐴 ∆𝑃 𝐹∙ 𝐶 𝜋𝑑
𝜋
𝑘

2∆𝑃
𝜌

𝑛𝐶 𝑑

𝐴 ∆𝑃 𝐹𝑘2∆𝑃𝜌5

EXPERIMENTAL VARIDATION OF DAMPING FORCE GENERATOR MODEL
Figure 6 shows the result of calculation of the variable damping force generator model. The proposed model can
express the nonlinear characteristics of the damping force generator for each controlled air pressure. This model
fits actual damper behavior well.

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Figure 6. Comparison of measured damping force characteristics and damping force characteristic model for damper rod
extension（Damper force and velocity are converted to pressure and flow rates）

CONCLUSION
In this paper, a model of the nonlinear damping force characteristics of an active suspension for compact cars isconstructed, which consists of an air spring using a pressure-resistant thin wall metal bellows and a variable
damping force damper. Tools simulating the behavior of vehicles equipped with conventional suspension are
widespread at the design stage. By incorporating the proposed model of damping force generator into the damper
section, it is expected to estimate the performance and vehicle behavior of the developed active suspension at thedesign stage.

REFERENCES
1. SATO, Y., MIYAZAWA, J., HIRAYAMA, T., SHIMBORI, T., Pressure Resistance Improvement of Metal Bellows for
Pneumatic Actuators, Proceedings of the 9th JFPS International Symposium on Fluid Power Matsue 2014, 2014, pp.284289.
2. OGAWA, T., SATO, Y., SHIMBORI, T., Air Spring Suspension for Compact Vehicle Posture Control using Metal
Bellows Actuator, Proceedings on Autumn Conference of JFPS 2017, 2017, pp. 95-97.(in Japanese)
3. McCLOY, D., MARTIN, H., Control of Fluid Power: Analysis and design 2nd edition, Ellis Horwood Limited, 1980,
pp.101-103.

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Power Absorber by Air Compression and Release Cycle with
Controller for Adjusting Air Release Timing to Obtain Variable
Absorbing Capacity (First Report – Basic Concept and Design)
Takashi SHIBAYAMA* , Toshinori FUJITA* and Sho UEYAMA**

Department of Advanced Machinery Engineering, Faculty of Engineering, Tokyo Denki University
5 Senju-Asahi-cho Adashi-ku, Tokyo, 120-8551 Japan(E-mail: t-shiba@f5.dion.ne.jp)

Abstract. To reduce CO2, fuel cell system for heavy duty trucks is under development popularly in the world.
For conventional heavy trucks with internal combustion engine (Hereafter ICE), power absorber by air
compression and release cycle is widely utilized to obtain effective braking force on long steep down hill road.
However, there is no ICE in fuel cell system. Regenerative brake is also effective for those heavy FCV trucks on
the condition that battery state of charge (hereafter SOC) is not full. If battery SOC becomes full, theregenerative brake could not be available. Some emergency power absorber is required for such severe situation.
In this paper, new power absorber by air compression and release cycle is proposed which includes controller
for adjusting air release timing to obtain variable capacity.
Keywords: Automobile, FCV, Brake, Energy absorber, Compressor

INTRODUCTION
The basic principle of proposed power absorber was reported at CTI Berlin conference in 2019[1]. Figure 1
shows the principle how resistance force generated. Air chamber is made by two different diameter pistons. The
two pistons are connected to reciprocating rod and the reciprocating rod is connected to rotating shaft similar as
crank system of the ICE.（In this figure, rotating shaft and crank system are not drawn） As shown in the
figure, air is in-taken at bottom dead center (Hereafter BDC), then air is compressed, and high resistance force isgenerated by high pressure. At top dead center (Hereafter TDC), the compressed air is released and exhausted.
After TDC, air is expanded and negative pressire is generated,and the resistance forece is generated again by thenegative pressure. By those resistance forces , power aborption is obtained. As shown in the figure, there is no
moving cams which are quite common for ICE.The moving pistons directly open and close the air inlet and air
outlet. Such simple system is aimed in this development. At low speed, isothermal calculation according to theformula “PV=const” can be used,. At high speed , adiabatic calculation can be used and compression pressure
will become higher. Here, isothermal calculation is applied to keep margin for designning large enough capacity
power aborption .

Figure 1. Basic principle of air compression and release cycle power absorber

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ADDITIONAL CONTROLLER FOR ADJUSTING AIR RELEASE TIMINGIn Figure 2 , new power absorber with controller for adjusting air release timing is shown. Additional controller
as shown in the figure 2 is applied. By the screw bolt and nut for adjusting air release timing, distance between
the two sliding O-rings becomes variable and adjustable that can be operated by a rotational actuator.

Figure 2. Improved power absorber with air release timing adjusting function.

SOME EXAMPLES OF AIR RELEASE TIMING ADJUSTING
To make the new system more comprehensive, let us see four examples. Figure 3 shows the example when theadjustment makes “ No air release” situation. Air outlet could not be opened even at TDC. Positive pressure
will remain, and power absorption is not high enough.

Figure 3. Adjustment Example-1

Air release adjustment makes ”No air release “ situation2

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Figure 4 shows the adjustment for maximum power absorption. By this adjustment, power absorber capacity
becomes maximum which is equivalent to Figure 1. Compressed air can be released and exhausted at TDC.
And after TDC, negative pressure during expansion process is generated. By both the positive pressure during
compression process and the negative pressure during expansion process generate resistance force against thereciprocating axis driving force. Then, maximum power absorption can be executed

Figure 4. Adjustment Example-2

Power absorber capacity is maximum, equivalent to figure 1

Figure 5 shows the one other example of adjustment for small capacity. Air will be released earlier than figure 4.
Because the air release timing is earlier, maximum chamber pressure is lowered and both positive pressure
duration length and negative pressure duration length are shortened, that leads to capacity reduction.

Figure 5. Adjustment Example-3

Power absorber capacity is small

And figure 6 shows the last example. By this adjustment, power absorber capacity becomes very small. If the air
release timing would be much earlier than this figure , air release can be started at BDC. By such operation, no
absorption situation can be also available.

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Figure 6. Adjustment Example-4

Powe absorber capacity is very small, almost always released

DESIGN FOR CONFIRMATION TESTING
To make confirmation testing, small desktop size power absorber has been designed and capacity estimation has
been executed. Figure 7 shows the result of the capacity estimation. As shown in Figure 7, power absorbercapacity can be enlarged by utilizing supercharger for inlet air. If inlet air pressure is larger than atmosphericpressure, pressure of compressed air becomes relatively large, and resistance force also becomes relatively large.
The result of planned confirmation testing will be reported at the next report.

Figure 7. Estimated capacity of the new power absorber prototype for desktop testing.

CONCLUSION
Power absorber by air compression and release cycle with controller for adjusting air release timing has been
investigated, and improved power absorber concept is established. Desktop size prototype will be tested and thetest result will be reported at the next report.

REFERENCE
1. SHIBAYAMA,T ,How to ensure safety of EV or FCV without reliable, helpful saving functions made by ICE, CTI
Berlin conference , 2019, Session G-5

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Clarification of Parameters and Development of a Method for
Estimating Loading Forces Acting on The Spool Valve of a
Hydraulically Controlled Automotive Transmission
Daisuke YANAGAWA*, Masahiro KOUYA*, Sho TOZUKA*,
Masaru SHIMADA**, Idris TENGKU*** and Naoki UEZONO*****

Department of hardware system development, **CAE group,
Department of experiment, **** Experiment Department No. 2
JATCO Ltd560-2 Okatsukoku, Atsugi, Kanagawa 243-0126, Japan(daisuke2_yanagawa@jatco.co.jp), (masahiro2_kouya@jatco.co.jp), (sho_tozuka@jatco.co.jp),(masaru_shimada@jatco.co.jp), (idris2_tengku@jatco.co.jp), (naoki_uezono@jatco.co.jp)***

Abstract. The control valve used in hydraulically controlled automotive transmissions contains a spool valve thatdrains the excess portion of the fluid flow rate while regulating the pressure. The complex flow around the spool
valve exerts various loading forces on the valve. These forces can delay valve response owing to increased sliding
resistance and also press the valve against the control valve body, giving rise to concern about valve body wear.
Although the forces acting on the valve can be calculated by fluid simulations, it has not been known whether thesimulation results are correct. Therefore, in this study a load cell was used to measure the loading forces directly.
The results confirmed that the tendencies of the calculated results obtained by fluid simulations agreed well withthose of the experimental data, thereby verifying the validity of the simulation results.
Keywords: Automatic transmission, Control valve spool, Loading force, Draft angle, Groove depth.

INTRODUCTION
An automotive transmission uses a control valve to supply fluid to multiple hydraulic circuits and to regulate thepressure. Located inside the control valve is a spool valve that drains the excess portion of the fluid flow rate
while adjusting the pressure. Loading forces induced by the fluid flow act on the spool valve during the draining
process and exert various effects on the valve.
However, heretofore it has been difficult to measure the loading forces acting on the spool valve, although they
have been calculated by fluid simulations. Consequently, it has not been known whether the simulation results
have been correct. For that reason, judgments about the feasibility of changing transmission specifications haveto be based on experimental evaluations of response, wear and other performance parameters. Because
specification changes require judgments based on experimentation, a lot of trial and error is involved, which canaffect the product development period.
In this work, an experimental setup was built that can directly measure the loading forces acting on the spool
valve. This paper presents the results measured with this setup, which confirm the validity of fluid simulations.

SPOOL VALVE STRUCTURE AND
CONFIRMATION OF DIMENSIONAL EFFECTS
Fig. 1 shows the structure of the spool valve for which experimental measurements were made in this study. As
fluid flows from upstream to downstream, the loading forces produced by the fluid act on the valve to increase
friction. That can cause a delay in responsiveness, which is a key characteristic for transmission control. In
addition, the loading forces press the spool valve against the control valve body in which it is housed, giving rise
to concern about valve body wear.
The experiments measured the effects of two dimensions, the draft angle and the groove depth. It was inferredfrom fluid simulation results that these two dimensions affect the loading forces.

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Figure 1 Loading forces acting on the spool valve and their effects

The draft angle refers to the angle of inclination of the valve body wall and is defined as shown in Fig. 2. This
inclination is provided to enable easy removal of the mold in the casting process, and the angle is retained by thevalve body. The groove depth is defined as shown in Fig. 3.

Figure 2 Definition of draft angle

Figure 3 Definition of groove depth

EXPERIMENTAL SETUP
It would be difficult to directly measure the loading force that the spool valve inside the control valve applies tothe latter valve body. However, the experimental setup was designed so that the force applied by the spool valve
can be measured by a load cell via a rod used as a probe. The position of the rod was determined fromsimulation results and by confirming the point from wear marks where the spool valve pressed against an actual
valve body. A hole was then drilled at that position for installing the rod. The rod passed through a jig to comein contact with the load cell installed above. This structure made it possible to measure the loading forces acting
on the spool valve using the load cell via the rod.
The structure of the experimental setup is shown in Fig. 4. The shape of the area around the spool valve in an
actual control valve was cut out in the jig, and the attachment parts shown in Fig. 5 enabled the dimensions
around the spool valve to be varied. This setup made it possible to determine how the loading forces wereaffected by the dimensions around the spool valve.
In addition, a screw was provided in the jig at the end of the spool valve for adjusting the amount of stroke fromoutside the jig so that the pressure would apply force to the rod at the proper stroke position during theexperiment.

Figure 4 Structure of jig, attachment parts and rod

Figure 5 Attachment parts

Fig. 6 shows the measured output of the load cell, representing the loading force applied to it when the upstreampressure was varied using the experimental setup described here. The specifications used in the experiment were2

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The 11th JFPS International Symposium on Fluid Power 2020a draft angle of 2.0 deg. and a groove depth of 11.5 mm. The vertical axes show the pressure and the loading
force applied to the load cell in relation to time on the horizontal axis. The red line is the upstream pressure andthe black line is the loading force.
The measured results show that the loading force rose along with the increase in the upstream pressure,indicating that the loading force followed the change in pressure. Presumably, this means that the flow rateincreased owing to the differential pressure between upstream and downstream, thereby increasing the loading
force produced by the fluid flow. These results confirmed that the experimental setup was capable of measuring
the loading force.

Figure 6 Output of load cell related pressure

SIMULATION MODEL
A 3D model was created of the jig, attachment parts and spool valve described above and used to conduct a fluid
simulation of the loading force that was produced by the hydrodynamic force and acted on the position of therod.
An example of the simulation model is shown in Fig. 7. The loading force simulated in this study acts in theannular clearance between the spool valve and the control valve body as illustrated in Fig. 8. Accordingly,
computational accuracy was improved by modeling the annular clearance with a finer mesh.
The simulation conditions were aligned with those of the experiment. In both cases, the upstream pressure was
5.7 MPa, the downstream pressure was 0.7 MPa, the fluid temperature was 50 °C, and the specification of thespool valve diameter was 17 mm.

Figure 7 Simulation model

Figure 8 Definition of clearance

COMPARISON OF EXPERIMENTAL AND SIMULATION RESULTS
The results obtained for the effect of the draft angle are shown in Fig. 9. The graph plots the loading force on thevertical axis as a function of the draft angle on the horizontal axis. The blue line is the experimental result andthe red line is the simulation result. A loading force of 6.3 N was measured under a condition of a draft angle of
0 deg. It is seen that the loading force increased with a larger draft angle, and a value of 19.3 N was measured at
an angle of 2.0 deg. The results of the fluid simulation conducted under the same conditions show a loading
force of 14.8 N under a condition of a draft angle of 0 deg. and a value of 24.6 N at an angle of 2.0 deg.
Although there is some discrepancy in absolute values between the experimental and simulation results, bothsets of data show the same tendencies and effects on the loading force, thus confirming the validity of thesimulation.

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Figure9 Loading force as a function of draft angle

It is assumed that the reason for the effect of the draft angle on the loading force is that it produces a pressure
unbalance between the circuit side and the back side, thereby applying force to the circuit side.
As the draft angle increases, a difference occurs in the size of the valve opening between the circuit side and theback side, with the opening becoming larger on the circuit side (Fig. 10). As a result, because the flow rate on
the circuit side increases, the flow velocity becomes faster than on the back side. The pressure on the circuit side
thus decreases, resulting in a differential pressure. The pressure around the spool valve on the circuit side
declines locally to produce the pressure unbalance shown in Fig. 11.

Figure 10 Effect of draft angle

Figure 11 Simulation results for draft angle

Fig. 12 shows the results obtained for the effect of the groove depth. The loading force is plotted on the verticalaxis as a function of the groove depth on the horizontal axis. The blue line is the experimental result and the redline is the simulation result. A loading force of 11.2 N was measured at a groove depth of 3.0 mm. The loading
force decreased as the groove depth was increased, declining to 3.3 N at a groove depth of 11.5 mm. The
simulation results show a loading force of 9.1 N at a groove depth of 3.0 mm and a value of 3.8 N at a groove
depth of 11. 5 mm.
Although there is some discrepancy between the experimental and simulation results, both sets of data show thesame tendencies, as was seen for the draft angle.

Figure 12 Loading force as a function of groove depth

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The action of the following two factors can be considered to explain the reason for the effect of the groove depthon the loading force. (1) The pressure rise induced by the collision of the fluid with the spool valve applies force
to the circuit side. (2) The pressure gradient between the back side and the circuit side raises the pressure on theback side, thereby applying pressure to the circuit side.
As shown in Fig. 13, the first factor concerning the collision of the fluid with the spool valve occurs as the fluidin the grooves on the back side flows toward the circuit side (Fig. 14). In the case of a deep groove depth, theflow spreads out before the collision occurs so the effect of the collision is smaller. In contrast, in the case of ashallow groove depth, the effect of the collision is larger so a pressure rise is observed.

Figure 13 Simulation results for groove depth

Figure 14 Effect of groove depth

As shown in Fig. 15, the second factor concerning the pressure gradient occurs in the direction of the flow of thefluid between the spool valve and the valve body when the fluid in the grooves on the back side flows toward
the circuit side. This is more pronounced for a shallow groove depth because the flow passage is narrower. The
influence is small in the case of a deep groove depth (Fig. 16). As a result, the pressure on the back side rises
compared with that on the circuit side, thus increasing the loading force.

Figure 15 Simulation results for groove depth in sect A-A

Figure 16 Effect of pressure drop

CONCLUSION(1) Using a jig cut out in the shape of the area around the spool valve, a rod as a probe and a load cell, theloading forces acting on the valve were measured directly, which has been difficult to do heretofore.
(2) A comparison of the experimental and simulation results revealed that there was some discrepancy in theabsolute values, but both sets of data agreed well regarding the tendencies of the dimensional effects on theloading forces. This confirmed the validity of the simulation results.
(3) This has now made it possible to make design judgments of transmission specifications at the development
stage before conducting physical tests.

REFERENCES
1. Kenji Sakakibara, ”Development of a hydraulic control system for a new JATCO wide-range CVT”, 10th International
CTI Symposium USA. Automotive Transmissions, HEV and EV Drives, USA, 2016, B2_Sakakibara_Jatco_paper, pp.
5-6.
2. Masaru Shimada, ”Application of CFD to develop a control valve spool featuring reduced fluid force”, 2017 JSAE
Congress (Autumn), No. 20176045 (2017).

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Organized Session | Organized session

Functional Fluid[OS2-01] A Study on a Mathematical Model for AC Electroosmosis Micropump
Yo Makita1, 〇Kazuhiro Yoshida1, Sang In Eom1, Joon-wan Kim1 （1.Tokyo Institute ofTechnology）[OS2-02] Multi-Layered Disk Type of Electro-Rheological Braking Device for Small
Mobile Robot〇Takanori Togawa1, Yuta Sato1, Yutaka Tanaka1, Jinghui Peng2 （1.Hosei University, 2.Harbin
Institute of Technology）[OS2-03] Lightweight &Twin-driven MR Fluid Actuator for Haptic Devices〇Tetsumasa Takano1, Asaka Ikeda2, Akinori Yamaguchi1, Isao Abe2, Takehito Kikuchi2（1.Graduate School of Engineering, Oita University, 2.Faculty of Science and Technology, Oita
University）[OS2-04] Investigation of Constant Velocity Motion with Physical Interaction System
for Long- Term Stay in Microgravity Space〇Tetsuhito Fujita1, Katsuki Machida1, Yusuke Shimoda1, Manabu Okui1, Rie Nishihama2, Taro
Nakamura1 （1.Chuo University, 2.Research and Development Initiative of Chuo University）[OS2-05] Manufacturing and Evaluation of Micro Electrohydrodynamic Pumps withDifferent Scales and Similar Dimensions〇Kotaro Okada1, Masahito Nishikawara1,2, Shunichi Naito1, Hideki Yanada1, Hiroshi Yokoyama1（1.Toyohashi University of Technology, 2.Worcester Polytechnic Institute）[OS2-06] Study on a Low Pressure and Flowrate Driving of Micro Leg Joints for Soft
RobotsKoji Michishita1, Kazuhiro Yoshida1, 〇Joon-wan Kim1 （1.Tokyo Institute of Technology）©The Japan Fluid Power System Society

OS2-01

The 11th JFPS International Symposium on Fluid Power 2020

A Study on a Mathematical Model for
AC Electroosmosis Micropump
Yo MAKITA*, Kazuhiro YOSHIDA**, Sang In EOM** and Joon-wan KIM***

School of Engineering, Tokyo Institute of Technology**
FIRST, IIR, Tokyo Institute of Technology4259-R2-42 Nagatsuta-cho, Midori-ku, Yokohama, 226-8503 Japan(E-mail: yoshida@pi.titech.ac.jp)

Abstract. An AC electroosmosis (ACEO) micropump has been required for lab-on-a-chip to generate a water flow
with a miniaturizable simple structure without mechanical moving parts. For design of ACEO devices, an ACEO
mathematical model is needed, however, the previous model has large discrepancy between the simulated andmeasured velocities when physical parameters are used. To overcome the problem, in this study, a new mathematicalmodel was proposed considering the forming time and the nonlinear capacitance of the electric double layer andFEM simulations were performed for a simple ACEO convection generator. The device was fabricated using a
MEMS fabrication process and its flow velocities were measured. Through comparison with the simulated values,
the effectiveness of the proposed mathematical model was confirmed.
Keywords: AC electroosmosis (ACEO), Mathematical model, Micropump, MEMS, Functional fluid

INTRODUCTION
For lab-on-a-chip, a micropump using AC electroosmosis (ACEO) has been required [1], because it cangenerate a water flow with simple fixed electrodes. A mathematical model of ACEO has been establishedtheoretically [1, 2], which can estimate the relative variation of the flow velocities, however, the absolute value
is far larger than the measured value. In this study, we proposed a mathematical model considering the forming
time and the nonlinear capacitance of the electric double layer (EDL). Then we confirmed its validity throughcomparison between the simulated and the measure flow velocities

PROPOSAL OF MATHEMATICAL MODELAs shown in Fig. 1, when a container having a pair of plate electrodes on the bottom is filled with liquid likewater and applied AC voltage to, charged particles are gathered on the electrode and an EDL is formed. The
electric field acts on the charged particles and generates slip velocities on the electrodes. When the polarity of
the applied voltage is changed, since the polarity of the electric field is also changed, the direction of the slip
velocity is unchanged and convection is generated. This phenomenon is called the ACEO.
In the previous mathematical model of the ACEO [1], the EDL was modelled as a capacitor with a thickness of
the Debye length λD and the electric field was analyzed. The zeta potential ζ and the electric field component Etparallel to the electrode were derived. The slip velocity was calculated by the Helmholz-Smoluchowski equation:ε(1)u s = − ς Et ,μwhere, ε and μ are the permittivity and the viscosity of the liquid and the overbar means the time-averaged value.
However, the slip velocity calculated with physical properties was 100 times higher than the measured value,
because a delay due to the EDL formation could not be ignored. Hence, in this study, we proposed a newmathematical model considering the forming time and the nonlinear capacitance of the EDL. In this model, the(a) nT ≤ t < (n+1/2)T(b) (n+1/2)T ≤ t < (n+1)T

Figure 1. Working principle of AC electroosmosis (ACEO) (T = 1/f, n = 0, ±1, ±2, …)

1

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The 11th JFPS International Symposium on Fluid Power 2020(a) Physical model(b) Equivalent circuit

Figure 2. Proposed model of ACEO convection generator(a) Electric field strength(b) Velocity field

Figure 3. Simulation results of ACEO convection generator

rectangular parallelepiped bulk liquid with a cross section A and length L is modelled as a parallel connection ofan electric resistance Rbl = ρL/A and a capacitance Cbl = εA/L as shown in Fig. 2, where ρ is a specific resistanceof the liquid. In addition, the capacitance CEDL of the EDL and the zeta potential ζ are calculated by: z eς εS(2)cosh  i  , ς = 0.01V ,CEDL =λD 2kTa 
where, zi is the valence of the charged particles, e is the quantum of electricity, k is the Bolzmann constant, Ta is
the absolute temperature, and V is the applied voltage amplitude. As a result, the slip velocity is calculated as:ε(3)u s = − ς Et cos θ ,μwhere, the phase θ shows the forming time of the EDL.

SIMULATIONS AND EXPERIMENTS
The validity of the proposed mathematical model was examined by FEM simulations and experiments. The
ACEO convection generator shown in Fig. 1 was investigated. The electrode width and interval were 200 μm
and the 200 μm, respectively, and the diameter and the depth of the cylinder container were 6 mm and 500 μm,
respectively. The liquid was water and the simulations were performed using COMSOL MultiphysicsTM.
Figures 3(a) and (b) show the simulated electric field and the velocity. The convection flows could be generated.
Figure 4 left shows the fabricated ACEO convection generator. A 0.5 mm thick silicone plate with a 6 mmdiameter hole was attached to a glass wafer with a pair of electrodes. The electrode part was fabricated by a
MEMS fabrication process shown in Figs. 4 (a) to (f) as follows: (a) Ti/Au deposition on a glass wafer, (b)positive resist S-1818 (Rohm and Haas Electronic Materials) spin coating, (c) S-1818 patterning, (d) Au etching.
(e) S-1818 removal, and (f) Ti etching. The device could be successfully fabricated as shown in Fig. 4 left.
Figure 5(a) shows the simulated slip velocities on the electrode. The position x is defined as the distance fromthe electrode edge as shown in Fig. 3(a). The slip velocity is decreased and the polarity is changed about at x =150 μm, which is a stagnation point. Figure 5 (b) shows the top view of the fabricated device. The bright line
shows the stagnation line. The proposed model can predict the stagnation.
In the experiments, the flow velocities were measured by PTV (particle tracking velocimetry) using fluorescentparticles (Sigma Aldrich l9654: diameter 1.0 μm, density 1050 kg/m3, excitation wave length 470 nm,
fluorescence wave length 540 nm. Volume concentration 0.061 %). The movement of the particles wasmeasured with a microscope and analyzed, and the slip velocities were obtained.

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The 11th JFPS International Symposium on Fluid Power 2020

Figure 4. Fabrication process of ACEO convection generator(a) Simulated slip velocity(b) Measured stagnation

Figure 5. Simulated and measured results(a) For position on electrode at 7 Vpp(b) For applied voltage at 100 Hz(b) For applied voltage at 300 Hz

Figure 6. Simulated and measured slip velocities of ACEO convection generator

The simulation and experimental results were compared. Figure 6(a) shows the slip velocity distribution on theelectrode. The simulation results are shown by the dashed, solid and dotted lines. The experimental results areshown by the diamond, circle and square symbols. The applied voltage was 7 Vpp. The discrepancy was not sosmall, however, the same order results were obtained. Figures 6(b) and 6(c) shows the slip velocity at x = 30 μm
for the applied voltage. At the applied voltage frequency 100 Hz, good agreement was confirmed (Fig. 6(b)). At
300 Hz, better agreement was obtained than that of the previous model (Fig. 6(c)).

CONCLUSIONSIn this paper, we proposed a new mathematical model of ACEO considering the forming time and the nonlinear
capacitance of the EDL and performed the FEM simulations. An ACEO convection generator was fabricated by
the MEMS technologies and the slip velocities were measured. As a result of comparisons, the better agreementof the simulation results using the proposed mathematical model was confirmed.

REFERENCES
1. SONI, G., Nonlinear phenomena in induced charge electroosmosis, Dissertation of University of California, Santa
Barbara, (2008).
2. YOSHIDA, K., SATO, T., EOM, S. I., KIM, J.-W., and YOKOTA, S., A Study on an AC Electroosmotic Micropump
Using a Square Pole - Slit Electrode Array, Sensors and Actuators A, 2017, 265, pp. 152-160.

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The 11th JFPS International Symposium on Fluid Power 2020

Multi-Layered Disk Type of Electro-Rheological Braking Device
for Small Mobile Robot
Takanori TOGAWA*, Yuta SATO*, Yutaka TANAKA**, Jinghui PENG **** Graduate School of Engineering and Design, Hosei University
2-33 Ichigaya-Tamachi, Shinjuku-ku, Tokyo, 162-0843 Japan(E-mail: togawa.takanori@gmail.com)** Faculty of Engineering and Design, Hosei University(E-mail: y_tanaka@hosei.ac.jp)*** Harbin Institute of Technology
Harbin, China, 150001(E-mail: pengjinghui009@gmail.com)

Abstract. In the micromouse competition, small autonomous mobile robots run at high speed in mazes. The robot
turns a corner at a steep angle in this competition. However, the recent competition is marked by speeding up. Thereis a limit to driving performance with motor current control braking. Electro-rheological (ER) fluid actuators have thepotential of soft brake function for the small autonomous mobile robot because the ER fluid can quickly control
changes in viscosity by changing the electric field. In this research the multi-layered disk type of mechanical brake
device using the ER fluid for the micromouse was designed and developed. The performance of the ER braking device
was verified through numerical simulation and experiments. Characteristics of on-board power supply for small brakemodels using the ER fluid was also numerically simulated. A drive electric circuit for boosting applied voltage to the
ER fluid and discharging stored electric charge due to high electric voltage for the small autonomous mobile robot
was proposed and numerically simulated.
Keywords: Electro-rheological fluid, Braking device, DC motor, Micromouse, Small autonomous mobile robot

INTRODUCTION
Micromouse competition is an event for self-made maze-solving autonomous small robot to compete intelligenceand performance [1]. The micromouse as shown in Figure 1 must conform to certain size limitations and needs tosolve the maze without outside human intervention. This robot event is held worldwide, especially popular in the
U.K., U.S.A., Japan, Singapore, India and South Korea. In this competition, it is required to skillfully control theacceleration and deceleration of a small robot based on its course. However, the acceleration/deceleration control
using the electric current control of the robots’ mounted motors limits the driving performance improvement. In
addition, it is necessary to mount mechanical components in a limited space, and the miniaturization of eachelement is an important issue. Inspired by the All-Japan Micromouse Competition, the newly designed small scale
soft brake aims to promote the driving performance of the micromouse.
Electro-rheological fluids (ER fluids) have characteristics that the viscosity changes due to the change of theelectric field between the electrodes, and the viscosity changes reversibly in a short time [2]. Since the viscosity
characteristics of the fluid can be controlled by changing the electric field, a braking device suitable forminiaturization with a simple structure without mechanical friction can be realized [3].
Our project team has been proposed and designed a new kind of small-scale soft braking device using ER fluid[4][5]. The innovation of the developed ER brake is that it is easier to miniaturize than mechanical braking(a) Micromouse(b) Micromouse in a maze
Figure 1. Typical small size of a micromouse and a maze1

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The 11th JFPS International Symposium on Fluid Power 2020mechanisms. And because it has less electrical energy, it is environmentally friendly. In addition, the braking
performance is improved compared to the electric motor alone, and the controllability and quick response ofelectric operation is also excellent for autonomous mobile robot. Optimal shape parameter of the disk type of thebraking device installed in the small mobile robot were designed and the dynamic characteristics of the braking
device were numerically simulated [6].
In this paper, a prototype model of the multi-layered disk type of the small ER braking device was fabricated andthe performance characteristics of the braking device were experimentally and numerically verified.
Characteristics of on-board power supply for the small braking device using the ER fluid were also numericallysimulated. A drive electric circuit for boosting applied voltage to the ER fluid and discharging stored electriccharge due to high electric voltage for the small mobile robot was numerically simulated.

ER FLUID AND ER BRAKING DEVICE
Figure 2 shows a configuration of the multi-layered disk type of the ER braking device for small mobile robots
such as micromouse. The cylindrical housing body has a diameter of 20.8 mm and a length of 32.0 mm. Threerotating disks of the positive electrodes are supported by a rotating center shaft and bearings. Four disk-shaped
negative electrodes are fixed by the cylindrical housing body. The thickness of the electrode is 0.5 mm and madeof copper. Gaps of the negative and positive electrodes are kept by conductive spacers at a constant value of 1
mm. The ER fluid is filled between the electrodes. The number of layers between the electrodes was selected as
six layers for the simulation of design.

Figure 2. Configuration of ER braking device

Figure 3. Inner structure view of the prototype ER braking device

Figure 3 shows the internal structure of the designed and prototyped ER braking device. The cylindrical housing
was manufactured as a half-split model and made of polyacetal (POM), which has high mechanical strength and2

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The 11th JFPS International Symposium on Fluid Power 2020

electrical insulating in engineering plastics. The disk electrodes, the shafts, and the spacers was made of
conductive metal copper and assembled inside the housing.
The characteristics of the practical ER fluid can be described using the Bingham model as shown in the following
equation:
𝜏 𝐸

𝜏

𝜇 𝛾,

1

where  is the shear stress, 𝛾 is the shear rate, E is the strength of the electric field between electrodes, 0(E) is the
dynamic yield stress varying with the strength of the electric field E, and 0 is the base viscosity of the ER fluid
without electric field. The dynamic yield stress 0(E) as a function of the applied electric field E and the baseviscosity 0 is given from preliminary experiments as following equations:
317.5 𝐸 .

𝜏 𝐸

0.225

𝜇

Pa ,

2

Pa ∙ s .

3

Figure 4 shows the typical relationship between the strength of electric field and the shear stress of the particle
type of the practical ER fluid under the shear rate of 400 s-1. As the strength of electric field increases, the sharestress increases like an exponential function. By using Eq. (2) and Eq. (3), the approximate curve in Figure 3 could
be obtained.

Figure 4. Shear stress vs. strength of the electric field for the ER fluid under the shear rate of 400 s-1

The braking torque TB can be obtained through the surface integral of the radius r as following equation [5]:
𝑁

𝑇

𝜏 𝐸

2𝜋𝑟

𝜇

𝑟𝜔
𝑑𝑟 ,
ℎ

4

where N is the number of layers, d1 and d2 denote the effective outer and inner diameters of the electrode disks, his the gap between each pair of positive and negative electrode disks, and  is the rotational angular velocity.
From Eq. (4), the braking torque of the 6-layer ER braking device shown in Figure 1 is given by the following
equation:
𝑇

𝜋𝜏 𝐸
2

𝑑

𝑑

3𝜋𝜇 𝜔
𝑑
16ℎ

𝑑

.

5

The rotational motion equation is expressed by the following differential equation:
𝐽

𝑑𝜔
𝑑𝑡

𝐾 𝜔

𝜔

𝑇 ,

6

where J is the moment of inertia for the entire rotating system, Kp (= 3.6 10-6 N m s/rad) is a coefficient
determined by the load characteristics of the DC motor, and 0 is the initial angular velocity of the system. The

3

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The 11th JFPS International Symposium on Fluid Power 2020

dynamic response of the rotating system with the DC motor and the ER braking device can be calculated by theEq. (5) and Eq. (6).

EXPERIMENTS
Figure 5 shows the configuration of the used equipment in the experiments. Brake characteristics wereexperimentally measured using the configuration. In the experiment, the DC motor was initially rotating in a
steady state with an angular velocity 0 (= 500 rad/s). Next, a voltage was applied to the ER braking device, thechange in angular velocity was measured by the encoder of the DC motor, and the measured value was stored in
the PC through the microcomputer. In these experiments, after starting the ER braking operation, the DC motor
has continued to generate torque and the electric short brake was not working.

Figure 5. Configuration of experimental setup

Figure 6 shows the experimental and simulation results of the transient response in the rotational angular velocity
when the ER braking device is operated. The applied voltage to the ER braking device is 1 kV, 2 kV, and 3 kV,
respectively. It can be seen that the ER braking device reduces the rotational angular velocity of the DC motor.
There is a slight difference between the experiment and the simulation in the transient state. On the other hand,the final steady-state values of the experiment and simulation were almost the same under the three conditions.

Figure 6. Comparison between the experiment and simulation for angular velocity change of the rotating system

SIMULATION OF DRIVE ELECTRIC CIRCUIT FOR ER BRAKING DEVICE
The power source mounted on the small autonomous mobile robot called the micromouse is two 3.7 volts lithium
polymer secondary batteries. A high voltage must be applied between the electrodes to operate the ER braking
device. Therefore, it is necessary to design an electric circuit for driving the ER braking device that boosts the4

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The 11th JFPS International Symposium on Fluid Power 2020voltage of the secondary battery to a high voltage. On the other hand, the condenser of the boosting electric circuitand the ER braking device have a capacitance function. When the micromouse accelerates, it is necessary toquickly discharge the electric charge accumulated by braking mode.
Figure 7 shows the simulation model of ideal boost and discharge electric circuit for the circuit simulator,
commercial software LTspice. The voltage of the secondary battery is switched by the power MOSFET S1 on the
primary side of the power transformer. The voltage boosted by the power transformer is rectified and smoothedby the diodes D1, D2 and capacitor C1 on the secondary side. As a result, the capacitors are sequentially charged
and a high voltage is applied to the ER braking device. Discharge is carried out by switching the electric resistanceR1 and the transistor S2.

Figure 7. Simulation model of the electrical circuit simulator LTspice
3.5

Output voltage [kV]

3.0
2.5

Vout

2.0
1.5
1.0
0.5
0.0
0

2

4

6

8

10

Time [ms]Figure 8. Simulation results of the booster circuit for the ER braking device

Output voltage [kV]

3.5

Vout

3.0
2.5
2.0
1.5
1.0
0.5
0.0
0

20

40

60

80

100

Time [ms]Figure 9. Simulation results of discharge circuit for the ER braking device
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The simulation result of the boosted voltage change is presented in Figure 8. The voltage of the secondary battery
was set to 7.4 V, and the switch S1 was switched at 10 microsecond intervals. It was confirmed by simulation thatthe output voltage Vout can be quickly boosted to 3 kV in 1.5 ms.
Figure 9 shows the simulation results of the discharge circuit. The switch S2 was switched at 10 ms intervals. It
was confirmed by simulation that the voltage of the ER braking device boosted to 3 kV instantly dropped to 0 V.

CONCLUSIONSIn this paper the multi-layered disk type of the braking device installed in the small mobile robot was proposedand designed. The prototype model of the multi-layered disk type of the small ER braking device was fabricatedand the performance characteristics of the braking device were experimentally and numerically verified. The
characteristics of the on-board power supply for the small braking device using the ER fluid were numericallysimulated. The structure of the drive electric circuit for boosting applied voltage to the ER fluid and discharging
stored electric charge due to high electric voltage for the small mobile robot was proposed and the performanceof the drive electric circuit was numerically simulated.

ACKNOWLEDGMENTS
We express appreciation the technical support, Mr. Takuma TACHIBANA of the graduate student in the Graduate
School of Hosei University and Mr. Mizuki SHIGIHARA of the undergraduate student in Hosei University.

REFERENCES
1. All Japan Micromouse Contest 2020, https://www.ntf.or.jp/?p=340, (accessed on 05 May, 2021).
2. TIAN Y., LI C., ZHANG M., MENG Y., WEN Y., Transient response of an electrorheological fluid under square-wave
electric field excitation, Journal of Colloid and Interface Science, Vol. 288, 2005, pp.290-297.
3. TAN, K. P., STANWAY, R., BULLOUGH, W. A., Braking Responses of Inertia/Load by Using an Electro-Rheological(ER) Brake, Mechatronics, Vol.17, No.6, 2007, p.277-289.
4. FAN, X., SAKAMA, S., TOGAWA, T., TANAKA, Y., Design and Fabrication of ER Braking Device for Micromouse,
IEEE Proc. 7th Int. Conf. on Fluid Power and Mechatronics (FPM2015), 2015, p.729-733.
5. PENG, J., TOGAWA, T., TACHIBANA, T., TANAKA, Y., Numerical and Experimental Investigation on Braking
Characteristics of an Electro-Rheological (ER) Micro Brake, JFPS International Journal of Fluid Power System, Vol.11,
No.3, 2018, pp.124-129.
6. TOGAWA, T., TACHIBANA, T., TANAKA, Y., PENG, J., Hydro-Disk-Type of Electrorheological Brakes for Small
Mobile Robots, International Journal of Hydromechatronics, 2021, (in press).

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The 11th JFPS International Symposium on Fluid Power 2020

Lightweight & Twin-driven MR Fluid Actuator for Haptic Devices
Tetsumasa TAKANO*, Asaka IKEDA**, Akinori YAMAGUCHI*,
Isao ABE** and Takehito KIKUCHI***

Graduate School of Engineering, Oita University
Dannoharu 700, Oita 870-1192, Japan**
Faculty of Science and Technology, Oita University
Dannoharu 700, Oita 870-1192, Japan(E-mail: t-kikuchi@oita-u.ac.jp)
Abstract. We proposed a twin-driven MRF actuator (TD-MRA) using two multi-layered disc-type MR fluid clutches
for haptics. We have developed two types of the TD-MRA and compared their torque performances. In our previous
works, we designed a 0.6 Nm-class TD-MRA for core components of the haptic interface. However, the size andweight were not suitable for the compact unit. On the other hand, the controllable torque was redundant. From thesereasons, we redesigned the TD-MRA, and a 0.3 Nm-class device was developed and evaluated in this study. The
specification comparison and results of the mechanical tests were reported in this article.
Keywords: Magnetorheological fluid; haptic device; teleoperation; MR fluid actuator.

INTRODUCTION
Bilateral control concept [1] was proposed as a controlling method for high-fidelity teleoperation systems, e.g.
endoscopic surgical robots [2]. However, this concept has not been sufficiently utilized in real systems. For
example, owing to the limitations of the present actuator and sensor technology, the da Vinci Surgical System®
[3] mainly uses a unilateral controller and, hence, there is less haptic feedback for surgeons. By adding a feedback
loop, the machine can support human skills with gravity compensation and friction compensation. In addition, it
provides a better haptic perception to the operator while performing a remote operation task. This function is very
useful for enhancing operational accuracy and reducing psychological stress from the human operator in very
sensitive operation.
Therefore, establishing a high-performance force feedback system for remote-controllable robot systems isnecessary to enhance operational accuracy and reduce operators’ psychological stresses. Thus, we have developeda twin-driven MR Fluid actuator (TD-MRA) that can control fine haptic feedback [4]. In this actuator, two highperformance MR fluid clutches [5] are used for torque transmission. Fauteux et al. [6] had developed a twin driven
MRF actuator using planetary gears, and Song et al. [7] used the same mechanism for a haptic master. The twindriven mechanism has been used to reduce the basic friction of the actuators, and its low friction characteristics
are suitable for haptic devices. However, it is difficult to balance the frictional (and viscous) torque of the normal
and reversed rotations of the planetary-gear-based mechanism with the MRF device because of the difference inrotational velocities. Hence, we proposed the TD-MRA using two multi-layered disc-type MR fluid clutches for
haptics and a flat gear mechanism.
In our previous works [4], we designed a 0.6 Nm-class TD-MRA for core components of the haptic interface.
However, the size and weight were not suitable for the compact unit. On the other hand, the controllable torquewas redundant. From these reasons, we redesigned the TD-MRA, and a 0.3 Nm-class device was developed andevaluated in this study. The specification comparison and results of the mechanical tests were reported in thisarticle.

LIGHTWEIGHT MR FLUID DEVICEWe developed some types of compact MR Fluid devices with the basic structure shown in Fig.1. The basic
structure is rotational symmetry, and the double bearings hold the rotational shaft. The multi-layered discs were
fixed alternately on the housing and the core. The magnetic wire is winded and embedded in the magnetic core.
The discs and the magnetic core are made of iron, while the other parts are made of non-magnetic materials. An
oil seal is an important component of an MRF device. We used a single lip seal made of fluorocarbon polymer(#31612A, Starlite Co., Ltd.). Figure 2 shows the 0.3 Nm-class MR Fluid device. Table 1 shows the comparison

1

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The 11th JFPS International Symposium on Fluid Power 2020of the specifications for two devices. The 0.3 Nm-class has halves performances and size of the 0.6 Nm-class foralmost parameters. The optimal design method was developed in the previous works [5]. We can select optimallydesigned torque controllable devices on the torque requirements of application.

Figure 1. Basic structure

Figure 2. 0.3 Nm-class MR fluid device

Table 1. Specifications of the MR fluid device
Device type
0.6 Nm-class [4]0.3 Nm-class
Number of fluid layers
6
6
Thickness of MRF layer
0.2 mm
0.2 mmInner radius of disc12.5 mm
10 mm
Outer radius of disc
15 mm
12 mm
Turning number of the coil
270
154
Rated input current
1.0 A
1.8 A
Rated output torque0.62 Nm
0.26 Nm
Inertia of output parts
8.76×10-6 kgm2
2.47×10-6 kgm2
Mass
183 g
106 g

EVALUTION OF MR FLUID DEVICE
MethodThe torque responses of the device were measured with the experimental setup (Fig.3). The device was fixed on
an immovable plate and the output shaft was driven at constant speeds with an AC servomotor (SGMAV 08A,
Yasukawa) during tests. The torque was measured in real time with a torque sensor (UTMII-5Nm, UNIPULSE).
For static torque tests, rotational velocity was controlled at 10, 20, 30, 60, and 120 rpm, to clarify the effects ofrotational velocities. In addition, the electric current was controlled at 0.0 ~ 1.8 A for 10 s. The average torques
for each condition were calculated. For step tests, the velocity was controlled at 10 rpm, and the current wascontrolled stepwise with initial and final reference current, e.g., 0.0, 0.3, 0.6 and 0.9 A.

Figure 3. Testing machine for MR fluid device

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ResultExperimental results on the static torque are shown in Fig.4. The horizontal axis and vertical axis show the angularvelocity in rad/s, and the output torque in Nm, respectively. As shown in the figure, the relation between torqueand input current has nonlinearity, and the velocity-dependency of the torque is relatively smaller than the effectof the input current (MR effect).
Figure 5 shows a sample of the step responses. The horizontal axes show the time in seconds, and the vertical axes
show the output torque (left) and input current (right). Time constant, which is the time for the step response toreach 63.2% of its final value, is evaluated as a response time. The time constants were about 10 ms for theconditions in which the current starts from zero ampere. On the other case the time constants were less than 5 ms.

0.3

0.6 A
0.3 A

0.1

Current
Torque

0.5

0.1

0.0 A

0
0

0.2

1

10 rpmCurrent [A]

0.2

0.3
Torque [Nm]Torque [Nm]

1.8 A
1.2 A

0

4
8
12
16
Angular velocity [rad/s]

2.95

Figure 4. Static torque

0
3

3.05

3.1

Time [s]Figure 5. Step response

TWIN-DRIVEN MR ACTUATOR
The basic idea for the TD-MRA was reported in our previous study [4]. A parallel linkage mechanism combines
the output shafts of two MR fluid devices to make two directional controllable torques with a one-way motor anda pair of flat gears. To improve the back drivability, the MR fluid device was used as a high-quality torquecontrollable clutch. In addition, the rotational direction of the input motor was not varied. Two directional output
torques were controlled with two MR fluid devices with high torque / inertia ratio. A pair of 0.3 Nm-class MR
devices was installed in this structure (Fig.6). Its specifications were listed in Table 2.

Figure 6. TD-MRA with 0.3 Nm-class MR fluid device

Device type
Inertia of output parts
Size (L×W×H)
Mass

Table 2. Specifications of the TD-MRA
0.6 Nm-class [4]4.73×10-5 kgm2
353×105×85
1.97 kg

0.3 Nm-class
2.09×10-5 kgm2
260×100×80
1.28 kg3

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CONCLUSION
We proposed a twin-driven MRF actuator (TD-MRA) using two multi-layered disc-type MR fluid clutches for
haptics. In our previous works, we designed a 0.6 Nm-class TD-MRA for core components of the haptic interface.
To reduce the size and weight, we developed 0.3 Nm-class device and installed in the twin-driven mechanism.
The performance of the device was evaluated and showed quick responses. As the next stage, we apply this device
as core components of the haptic device.

ACKNOWLEDGMENTS
This research was partially supported by Adaptable and Seamless Technology transfer Program through targetdriven R&D of the Japan Science and Technology Agency (JPMJTM20HM), Innovative Science and Technology
Initiative for Security, ATLA, Japan (JPJ004596), and Grant-in-Aid for Scientific Research of Japan(JP17H03404).

REFERENCES
1. SICILIANO B and KHATIB O (2016) In: Handbook of Robotics (Second Edition, Section 43). Telerobotics. SpringerVerlag. Berlin Heidelberg, pp.1085–1108.
2. TAVAKOLI M, AZIMINEJAD A, PATEL R and MOALLEM M (2007) High-fidelity bilateral teleoperation systems andthe effect of multimodal haptics. IEEE Transactions On Systems, Man, And Cybernetics—Part B: Cybernetics 37 (6):1512–1528.
3. SIMOROV A, STEPHEN OTTE R, KOPIETZ C and OLEYNIKOV D (2012) Review of surgical robotics user interface:
what is the best way to control robotic surgery. Surgical Endoscopy 26: 2117–2125.
4. KIKUCHI T, ABE I, NAGATA T, YAMAGUCHI A, TAKANO T (2020), Twin-Driven Actuator with Multi-layered DiscMR Fluid Clutches for Haptics, Journal of Intelligent Material Systems and Structures, DOI: 10.1177/1045389X20943958
5. KIKUCHI T and ABE I (2019) Low inertia torque controllable device using magnetorheological fluid & umbrella-shaped
rotor. Engineering Research Express 1 (2): 02502. doi: https://dx.doi.org/10.1088/2631-8695/ab4ecf.
6. FAUTEUX P, LAURIA M, HEINTZ B and MICHAUD F (2010) Dual-differential rheological actuator for highperformance physical robotics interaction. IEEE Transaction on Robotics 26 (4): 607–618.
7. SONG B-K, OH J-S and CHOI S-B (2014) Design of new 4-DOF haptic master featuring magnetorheological fluid.
Advances in Mechanical Engineering, 843498.

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Investigation of Constant Velocity Motion with Physical Interaction
System for Long-Term Stay in Microgravity Space
Tetsuhito FUJITA*, Katsuki MACHIDA, Yusuke SHIMODA,
Manabu OKUI*, Rie NISHIHAMA, Taro NAKAMURA**

Department of Precision Mechanics, Faculty of Science and Engineering
Chuo University
1-13-27 Kasuga, Bunkyo-ku, Tokyo, 112-8551 Japan(E-mail: t_fujita@bio.mech.chuo-u.ac.jp)

Abstract. In this study, we developed a wearable device for training and rehabilitation using variable viscoelastic
joints and proposed a physical interaction system (PIS) using the device. In addition, the training effect of thedeveloped wearable variable viscoelastic joint unit and the elbow joint using this unit in constant velocity motion wasverified experimentally. The elbow joint was flexed and extended under different conditions, and the training effectwas verified by measuring the muscle potential. The results confirmed that the magnetorheological brake controlled
the elbow joint motion to a constant motion. Furthermore, the increase in muscle potential during the experimental
movements and the training effect of this unit during constant velocity movements were confirmed.
Keywords: Magnetorheological fluid brake, Artificial muscle, Training.

INTRODUCTION
A prolonged stay in microgravity or weightlessness causes muscle mass loss and decreased range of motion of
joints [1] [2]. Therefore, it is important to prevent these problems through training and rehabilitation during thestay. Currently, the International Space Station (ISS) uses advanced resistive exercise devices (ARED) and otherdevices to prevent muscle mass loss [3]. However, because of its large size and high weight, the ARED can only
be used in facilities such as the ISS, which are constructed through multiple round trips. In addition, the training
motions are limited, and rehabilitation to prevent a decrease in the joint range of motion cannot be performed.
Therefore, several studies have been conducted on various devices to maintain the physical functions of astronauts
in order to prevent deterioration of their physical functions [5] [6]. However, all of these devices use a motor as
the power source, which can be dangerous if the timing of the movement between the device and the person does
not match or if the device cannot be left attached and cannot be moved when it is not driven.
In this study, we aim to develop a wearable device that can be used for training and rehabilitation and to realize aphysical interaction system (hereinafter referred to as "PIS") using the device, with the objective of applying it tolong-term stays in microgravity and weightlessness, including the exploration of other planets. The concept of the
proposed PIS is shown in Figure 1. The proposed system has the characteristics of a wearable and unitary system.
The wearable design allows astronauts to train without any special preparation. In addition, the system can beused to provide resistance to daily movements to simulate gravity. One unit consists of two antagonisticallyarranged artificial muscles and a magnetorheological (MR) brake attached to the output shaft. The user canreconfigure the units to accommodate a variety of motions, such as bending and stretching with a wearable unit
or using it as an ARED [3] by fixing it on the floor. In this manner, training and rehabilitation can be conductedusing a single system, which reduces the system volume and weight, thus reducing costs. To realize the PIS, we
focus on the variable viscoelastic joint (hereinafter referred to as "the joint"), which was developed in a previous
study [7] [8]. This joint is composed of antagonistically arranged artificial muscles and an MR brake attached tothe drive shaft.. The joint is lightweight, has high power, and provides high backdrivability.
This study proposes and develops a wearable variable viscoelastic joint unit (hereinafter referred to as "the unit")
and verifies the training effect of isokinetic exercises on the upper limbs using the unit through experiments. The
target motions are flexion and extension of the elbow joint, and the training effect is verified by measuring themuscle potentials under each condition of changing the joint viscosity.

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Figure 1. Concept of the physical interaction system and image of the developed unit being freely reconfigured andattached to upper and lower limbs for training

CONCEPT OF PIS
The following functions are required for training and rehabilitation under microgravity.
(1) Support for various motions
In a microgravity environment, it is necessary to be able to respond to a variety of exercises rather than a fixed setof exercises because the entire body will lose its physical functions. This is evident from the functions of the
training devices currently in use.
(2) Lightweight and low volume
The space available at manned bases such as the ISS is limited, and the weight of the materials to be transported
from the ground to the manned bases greatly affects the cost. For these reasons, lightweight and low volume arenecessary features for the PIS.
(3) Application of load in zero gravity
In a microgravity environment, it is difficult to train and rehabilitate using one’s own weight or the weight ofother devices. Therefore, the PIS needs to function such that it can apply a load to the user even in a gravity-freeenvironment.
(4) Safety
Because the PIS envisioned in this study is a wearable device, safety must be ensured when it is worn by a user.
To satisfy these requirements, the PIS envisioned in this research (1) is capable of various motions depending on
the combination of units, (2) weighs less than 3 kg per unit and can be used in a wearable form, (3) controls therotational torque applied to the joints and uses it as a load, and (4) is a backdrivable and human-friendly systemthat imitates the principle of the human joint.

Classification of Training and Rehabilitation
Muscle contraction can be classified into eight types of movements, including movement and contraction
elements [9], as well as active and passive movements in which humans move proactively and passively inresponse to the movement of a device (Figure 2). The contraction exercises shown in Figure 2 can be performed
individually or in combination with each other to maintain muscle strength and demonstrate the effects ofrehabilitation. Therefore, the PIS proposed in this study aims to be a system that can perform all of these exercises
comprehensively. Hence, in this study, we first realize active movements in isokinetic contraction using viscous
control with this unit and then verify the training effect by measuring muscle potential.
StaticContraction
MuscleContraction

Isometric
Contraction

Isotonic
Contraction

Dynamic
Contraction
Isokinetic
Contraction

Passive
ActiveConcentricContraction
Eccentric
Contraction
ConcentricContraction
Eccentric
Contraction

Passive
ActivePassive

ActiveFigure 2. Classification of movements by type of contraction

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CONSTANT VELOCITY MOTION WITH VARIABLE VISCOELASTIC JOINTS
Variable Viscoelastic Joint System
In the PIS, high-precision position control is not important, but a structure that does not endanger the user isnecessary. In this study, we focused on the variable viscoelasticity property, which is the principle of the humanjoint.
The principle of the variable viscoelastic joint [7] [8] is shown in Figure 3. The joint is composed of two straightfiber-type pneumatic artificial muscles [10] (hereinafter referred to as "artificial muscles") and one MR brake,
which are antagonistically arranged to mimic the principle of the human joint. The variable elasticity of the joint
is achieved by applying air pressure to the artificial muscles. The angle and stiffness of the joint can be controlled
independently by controlling the air pressure on the artificial muscles. The MR brake is connected to the output
shaft, and the frictional torque of the MR brake is controlled based on the angular velocity of the output shaft toachieve variable viscosity of the joint.

Figure 3. Operating principle of the viscoelastic joint. This joint has antagonistic artificial muscles andmagnetorheological fluid brake. The joint stiffness and angle can be changed by applying air pressure to theartificial muscles.

Control Method of MR Brakes
Figure 4 shows an overview of the MR brake control. The control method of the joint viscosity in this joint is
described in Eq. (1), where τy is the frictional torque of the MR brake, θM is the measured joint angle, dθ/dt is theangular velocity, and Djd is the joint viscosity.
𝜏𝑦 = 𝐷jd

𝑑𝜃
𝑑𝑡

(1)

Figure 4. MR brake control overview

PROTOTYPE
Specifications of The Prototype
The following is a description of the prototype "Wearable Variable Viscoelastic Joint Unit" developed in thisstudy to apply variable viscoelastic joints to the PIS.
Figure 5 shows the actual photograph and characteristics of the MR brake (MR brake hollow type No. 505, ER
Tech, Japan) used in this unit. The MR brake used in this unit (Figure 5(a)) is 59 mm in diameter, weighsapproximately 280 g, and produces 6 Nm of brake torque when the maximum voltage of 8 V is applied.
Figure 5(b) shows the measured brake characteristics. To generate the command torque, an equation that linearly
approximates this current-torque relationship is used. In other words, Eq. (2) is used to generate τy. Because thedependence of the MR brake on the rotational speed is known to be low [7] [8], the rotational speed is notconsidered in the calculation of the command value.
𝜏y = 15.959𝐼 − 0.3283(2)

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Brake Torque [N]86y = 15.959x - 0.3283

4
2
0
-2

59 mm

0

0.1

0.2
0.3
Current value [A]

0.4

0.5

(a) Actual photograph of the MR brake used(b) Characteristic graph of the MR brake used
Figure 5. Structure and characteristics of the MR brake installed on the joint

DesignThe 3D CAD model of the wearable variable viscoelastic joint unit and the appearance of the manufactured unit
are shown in Figures 6 and 7, respectively. Two artificial muscles, a rotary encoder (MES-6-500-PC, Micro-tec,
Japan), a pneumatic valve (VDW20GA, SMC, Japan), and a pressure sensor (SEU11-6UA, PISCO, Japan) areinstalled in each unit. Each unit measures 400 mm in length and 120 mm in width and weighs approximately 2.0
kg. The force from the two artificial muscles is transmitted to the sprocket by a roller chain to generate torque.
The diameter of the sprocket used in this unit is 36.6 mm, and the range of motion of this unit is 135 deg. This
range of motion can be easily changed by changing the sprocket attached to the unit. In this experiment, only MR
brakes were used because the target was an active motion among constant velocity motions, and artificial muscles
were not used in the experiment.

Figure 6. 3D CAD model of the developed variable viscoelasticity unit
Artificial
Muscle

400 mm

110 mmPulleyMR Brake
120 mmRollerChain(a) Top View(b) Side View
Figure 7. Actual photograph of the developed unit

EXPERIMENTS WITH THE PROTOTYPE
Purpose of Experiment
The purpose of this experiment was to confirm the training and rehabilitation effects of isokinetic control by joint
viscosity and active isokinetic movement in the upper limb unit joints using an MR brake, a viscous element ofvariable viscoelastic joints. In addition, we verified the effect by measuring the muscle potential of the target area
during the flexion and extension movements of the arm.

Experimental Environment and Conditions

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The experimental environment is illustrated in Figure 8. The experiment was conducted under the condition ofviscous change generated by the MR brake. In the experiment, the subject sat on a chair and moved the handle of
the arm attached to the end of the apparatus. In this process, a digital force gauge (RZ-50, Aiko Electric, Japan)
was attached to the arm to measure the reaction force applied to the hand when the arm was moved. The subject
rotated the arm at the maximum speed at which the MR brake does not slip when a voltage is applied. The rotation
speed of the arm, the load applied to the hand, and the electromyography (hereinafter referred to as"EMG")potential of the upper limb were measured at this time. In addition, the subjects practiced each viscosity
thoroughly and understood the maximum speed in that viscosity before the experiment. In this experiment, thesubject was a healthy male (age, 24 years; height, 175 cm; weight, 55 kg). The upper limit of the voltage applied
to the MR brake was set to 8 V.
The target motion of this experiment was the flexion-extension motion of the arm, in which the elbow joint was
moved back and forth five times with an amplitude of 90 deg. The maximum flexion angle of the arm was set as
90 deg. In this experiment, the subject first held the handle attached to the arm with the forearm extended (thearm angle was 0 deg at this time). At the start of the experiment, the subject bent the forearm by 90 deg.
Immediately after the forearm was bent by 90 deg, the arm was extended by 90 deg. This movement was set as
one cycle, and the exercise was performed for five cycles in each trial. The results of three out of five trials wereused as the experimental results, and three of the five cycles of arm flexion and extension movements, excluding
the first and last cycles, were used. In addition, the load on the hand tip during flexion was measured in thisexperiment.
Muscle potentials were measured using the Trigno Wireless EMG system (DELSYS). The measurement points
were the biceps and triceps muscles. The measured EMG data were band-pass filtered (20–450 Hz), full-wave
rectified, and low-pass filtered (10 Hz).

Figure 8. Experimental environment

Results and Discussions
Table 1 shows the angular velocity of the arm and the load applied to the hand during motion with each commandviscosity.
The left column of Table 1 shows the joint angular velocity of the elbow during the experimental motion. The
vertical axis shows the angular velocity, and the horizontal axis shows the motion time. The solid line represents
the mean value of the movement speed, and the dashed line represents the standard deviation. In the graph, whenthe value is in the positive region, the human forearm moves in the flexion direction, and when it is in the negative
region, it moves in the extension direction. The graph shows that the standard deviation value does not deviatesignificantly from the mean value for any viscosity, and a constant motion is achieved. It indicates that theviscosity of the MR brake can control the joint angular velocity of the arm. When the command viscosity is 0.09
deg/s, which is the highest, the maximum angular velocity is 100 deg/s, and the minimum is -100 deg/s comparedwith the other viscosities. In the first half of the motion in the bending direction, the angular velocity is controlled
to be more constant than when the command viscosity is 0.03 and 0.06 deg/s.
The right column of Table 1 shows the load applied to the tip of the hand during the bending motion. The verticalaxis shows the load, and the horizontal axis shows the operation time. The graph shows that the higher the viscosity,
the lower the load applied to the tip, and the lower the viscosity, the higher the load. However, the lower theviscosity, the shorter the time in which the load is applied to the tip, and the more temporary the load is.
This is because in the case of low viscosity, the motion speed is high, and the inertia acting on the hand is larger,
resulting in a higher value of the load. In contrast, when the viscosity is high, the operating speed is low, and theoperating time is long, thus the time when the load is applied is long. It can also be confirmed that the values ofthe standard deviation of the load do not deviate significantly from the average value from the beginning to the
middle of the bending operation for all viscosities. However, at the end of the operation, the values of the standard
deviation of the load deviated from the mean value. This is because the subject knew that there was an extension

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movement after the flexion movement, so he stopped the movement and tried to move to the next movement,
which caused the load applied to the hand to be different for each movement.
Table 1. Angular velocity of the elbow joint and load applied to the hand during the experimental motion
Angular velocity [deg/s]
Torque [Nm]

0.03

90
70

Force [N]Angular Velocity [deg/s]

250
200
150
100
50
0
-50
-100
-150
-200

50
30
10

-10
-30
-50

0

0.2 0.4 0.6 0.8

10

1.2 1.4 1.6 1.8

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9Time [s]Time [s]
90
70

0.06

Force [N]Anglar Velocity [deg/s]

Viscous[Nm s/deg]

250
200
150
100
50
0
-50
-100
-150
-200

10

-30
-50

0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8Time [s]00.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9Time [s]

250

90

200

70

150

50

100

Force [N]Angular Velocity[deg/s]

30

-10

0

0.09

50

50
0

-50

30

10
-10

-100

-30

-150
-50

-200
0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 2.2 2.4
Time [s]

0

0.2

0.4

0.6

0.8

1

1.2

Time [s]Figures 9 and 10 show the values of EMG potential on the vertical axis and the motion cycle on the horizontal
axis for the biceps and triceps muscles, respectively. The orange, gray, and yellow lines represent the commanded
joint viscosity of 0.09 deg/s, 0.06 deg/s, and 0.03 deg/s, respectively, and the blue line represents the value when
the device is not attached. It can be observed that the biceps muscle is the flexion prime mover and therefore
increases in the first half of one cycle during flexion, while the triceps muscle is the extension prime mover andtherefore increases in the second half of one cycle during extension. At command viscosities of 0.03 and 0.06
deg/s, the EMG potentials of both biceps and triceps were higher than those without loading.
In this experiment, it was confirmed that there was a training effect when the command viscosity was 0.03 and0.06 deg/s since the muscle potential increased because of the viscosity change. When the command viscosity
was 0.09 deg/s, and when the device was not installed, the EMG potential values did not change compared withthe other command viscosities. In addition to the possibility that the inertia force generated during the movement
was canceled out, we believe that the subject may have become accustomed to the movement. He might have
moved in a way that did not greatly activate the muscles because the movement was performed at 0.09 deg/s
viscosity as the last change condition.
In addition, the contraction force of a muscle changes in inverse proportion to the contraction speed. This suggeststhat the maximum contractile force of the muscle is low for the command viscosity of 0.03 and 0.06 deg/s, whichhave high operation speeds, and that the value of the muscle potential is larger than that without the device because
it is necessary to activate many muscles to operate under the load applied by the viscosity. However, in the caseof the command viscosity of 0.09 deg/s, the motion speed was lower than in the other viscosity cases, and thecontraction force of the muscles could be higher, thus the number of active muscles was smaller, resulting in lower
EMG values. In addition, from the graph of the EMG potential of the triceps brachii muscle shown in Figure 10,

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the peak values when the command viscosity was 0.09 deg/s and when the device was not installed were different
from those when the command viscosity was 0.03 and 0.06 deg/s. This is because when the viscosity of the MR
brake is high, the load applied to the muscle force is changed by canceling out the inertia during operation.

Without

Vis0.09

Vis0.06

Vis0.03

70
60

EMG [μV]

50
40

30
20
10
0
024

1030

2036

3042

40 48

50 54

6060

70 66

80 72

9078

10084

Motion cycles [％]Figure 9. EMG potential values of the biceps brachii muscle

Without

Vis0.09

Vis0.06

Vis0.03

30
25

EMG [μV]

20
15
10
5
0
240

30
10

36
20

42
30

48
40

54
50

60

66
70

72
80

78
90

84
100

Motion cycles [％]Figure 10. EMG values of the triceps brachii muscle

CONCLUSION
In this study, we developed a wearable device for training and rehabilitation using variable viscoelastic joints andproposed a physical interaction system (PIS) using the device. In addition, the training effect of the developed
wearable variable viscoelastic joint unit and the elbow joint using this unit in constant velocity motion was verified
experimentally. The elbow joint was flexed and extended under different conditions, and the training effect wasverified by measuring the muscle potential. The results confirmed that the MR brake controlled the elbow joint
motion to a constant motion. Furthermore, the increase in muscle potential during the experimental movements
and the training effect of this unit during constant velocity movements were confirmed.
In the future, we aim to realize various motions with the device attached to a single joint of the lower limb.

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REFERENCES
1. OSHIMA, H., MIZUNO, K. and KAWASHIMA, S., Effects of Space Flight on Bone and Muscle, and Astronaut Exercise
Program, 2006, The Japanese Journal of Rehabilitation Medicine.
2. KAWASHIMA, S., MIZUNO, Y., WATANABE, Y., OSHIMA, H., Verification about effectiveness of resistance training
against changing muscular volume and strength. 2006, JAXA Special Publication: Scientific Report of International Bed
Rest Study, JAXA-SP-06-007, pp.34-68.
3. NASA，TEXAS INSTRUMENTS. (n.d.). ARED – RESISTIVE EXERCISE IN SPACE.
https://www.nasa.gov/pdf/553871main_AP_ST_Phys_ARED.pdf (accessed on 21 November, 2020, from
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5. NASA. GFE Exercise, https://ntrs.nasa.gov/api/citations/20170009178/downloads/20170009178.pdf6 (accessed on 21
November, 2020)
6. NASA, X1: A Robotic Exoskeleton for In-Space Countermeasures and Dynamometry.,
https://ntrs.nasa.gov/citations/20140000694 (accessed on 21 November, 2020)
7. Okui, M., Iikawa, S., Yamada, Y. and Nakamura, T., Fundamental characteristic of novel actuation system with variable
viscoelastic joints and magnetorheological clutches for human assistance, Journal of Intelligent Material System andStructures, 2018.
8. Okui, M., Iikawa, S., Yamada Y. and Nakamura, T. Variable viscoelastic joint system and its application to exoskeleton,
IEEE/RSJ International Conference on Intelligent Robots and Systems, pp. 3897-3902, 2017.
9. ICHIHASHI, N., Basic Knowledge of Strength Training - Factors Affecting Strength and Mechanism of Strength Increase-, 1997, Bulletin of the College of Medical Technology, Kyoto University. Supplement, Health and Human Sciences.
http://hdl.handle.net/2433/49548.
10. KOJIMA, A., OKUI, M., HISAMICHI, I., TSUJI, T., Nakamura, T., Straight-Fiber-Type Artificial Muscle Deformation
Under Pressurization. 2019, IEEE Robotics and Automation Letters.

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Manufacturing and Evaluation of Similar
Micro-Electrohydrodynamic Pumps with Different Scales
Kotaro OKADA*, Masahito NISHIKAWARA*, **, Shunichi NAITO*, Hideki YANADA*,
Hiroshi YOKOYAMA**

Department of Mechanical Engineering, Toyohashi University of Technology
1-1, Hibarigaoka, Tenpaku-cho, Toyohashi, Aichi, 441-8580 Japan(E-mail: Okada.kotaro.ts@tut.jp)**
Mechanical Engineering Department, Worcester Polytechnic Institute
100 Institute Road, Worcester, MA 01609 USA

Abstract. We investigated the effect of electrode scale on the performance of Electrohydrodynamic (EHD) conduction
pumps in this work. Because there are no moving parts in EHD pumps, they can be operated very quietly. Using
MEMS technology, we produced micro sized EHD pumps with varying scales and conducted experiments toinvestigate effect of electrodes scale on EHD pumps. Large pressure was generated with lower flow channels. EHD
simulation was conducted with COMSOL Multiphysics to investigate heterocharge layer structure on each scale. Whenthe space between the electrodes was smaller than 50 μm, the positive and negative heterocharge layers overlap. The
overlap is needed to be considered in the microscale electrode design.
Keywords: Electrohydrodynamics, EHD pump, MEMS, Scale effect

INTRODUCTION
Electrohydrodynamic (EHD) flow is generated without the use of moving mechanical parts and dielectric fluids
are usually used as the working fluid. The EHD flow is promising for pumps, actuators, and heat transport [1-4].
EHD pumps have low vibration and low noise, are easy to miniaturize, and are electronically controllable. Typicalphenomena that drive the dielectric fluids are ion drag and conduction. Conduction phenomenon is a flow
phenomenon driven by the Coulomb force acting on the electric charges generated by dissociation. In this paper,
we treat EHD pumps.
The electrodes widely used for EHD conduction pumps are flush electrode pair with different length. Thisgeometry generates a net flow in the direction to the longer electrode [5-7]. The advantages of this geometry arethat it can be used in series to obtain high pressures and can be placed as a part of other device with high flexibility.
EHD pumps with several microscale electrodes have been studied so far, as shown in Table 1, and the microscale
EHD pump characteristic and the effect of electrode asymmetry on the pump characteristic have been investigated[8-17]. Figure 1 shows the schematic of the EHD conduction pump of this research. The shape, dimensions of
the electrodes, L1, L3, and the electrode gap, L2, and the ratio of these dimensions were determined in a previous
research [18]. The ratio of L1, L2 and L3 is 1:1:3. The asymmetry of the electrodes generates a Coulomb force
unbalance pushing the dielectric working fluid from the narrow electrode to wide electrode [19].
Vázquez et al. [20] have done fundamental numerical simulations of EHD conduction pumping with two-parallel
electrodes and introduced two nondimensional numbers, C0, and β. C0 is called conduction number which is theratio of the distance that a dissociation charge migrates during relaxation time to the electrode scale, and β is theparameter that accounts for the electric field enhanced dissociation of dissociative molecules. They have shown
two operational regimes of the EHD conduction, called ohmic regime and saturation regime. In the ohmic regime(C0 >1), positive and negative charges are accumulated on the opposite polarity electrode, respectively, and thelayer of the charges with opposite polarity to each electrode is called heterocharge layer. Electric current is
proportional to the applied voltage and developed pressure increases with the square of applied voltage. In
saturation regime (C0 <1), on the other hand, as the heterocharge layer is thicker than the electrode spacing, bothheterocharge layers overlap. Michel et al. [21] showed that heterocharge layers overlap with asymmetric flush
electrodes of L2=127 μm in EHD conduction pumping used in a flow distribution control device.
Several microscale EHD pumps have been developed. However, The effect of electrode scale on the EHD pumpcharacteristic has not been clarified so far. Therefore, in this study, we use MEMS technology to produce similar
EHD pumps with different scales and we investigate the effects of the pump scale on EHD pump characteristics
experimentally and numerically.

1

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Figure 1. Schematic of EHD pump (not-to-scale).

Refs.

Table 1. EHD pumps dimensions in previous research [8-17].
L1 μm
L2
L3 μm
L4 μm
pair
Voltage Channel
0(narrow)μm(wide)
(pair
kV
width
n
(gap)
pitch)mm

Channelheightmm

Channellengthmm

M. K. Russel
2016 [8]

20

30

40

60

100

0.8

-

0.1

-

M.R. Pearson
2014 [9]

1220

970

4060

16510

9

5

57

57

152

M.R. Pearson
2013 [10]

127

127

381

1270

21

0.75

12.7

0.508

40

I. Kano 2013 [11]

120

60

360

480

20

0.5

21

0.1

16

M.R. Pearson
2011 [12]

51

51

178

794

10

0.8

12.7

0.254

11

Y. He 2010 [13]

20

20

20

80

-

0.2

-

0.1-0.15-

Y. Sakurai 2009
[14]

200

200

200

700

69

2

7

1

90

S. Yokota 2008
[15]

80

200

80

400

108

3

11

0.5

90

J. Darabi 2005
[16]

10

20

10

40

200

0.2

10

0.050

16

S. Moghaddam
2005 [17]

3

10

3

10

-

0.07

32

-

32

This study

5,
50,
500

5,
50,
500

15,
150,
1500

50,
500,
5000

200,
20,
2

0.05,
0.5,
5

18.7

0.02,
0.2,
2

15

2

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EXPERIMENTAL APPARATUS AND METHODEHD Pump and Design Electrodes
To generate the net Coulomb force acting on dissociated charges, the electrode configuration was asymmetric.
To investigate the effect of electrode scale and channel heights, MEMS fabrication technology was used to
produced similar EHD pumps with varying sizes. To examine the impact of channel height, all pumps have achannel height of h = 4L2 (see Figure 1) and pumps with L2 = 50 µm have three different channel heights, 1L2,
4L2, and 20L2. β and C0 are the nondimensional numbers used to define the effects of the scale on EHD pumps,
and they are described in the numerical model section. Because the channel length is the same, a pump withsmaller electrodes has more electrodes. A fluorinated fluid (Vertrel XF) is filled in the pump as the working fluid.
We produced EHD pumps using photolithography, a MEMS technology. For the electrode fabrication, we used a
Cr coated glass substrate. The thickness of the Cr coating was 100 nm and the resist (FMR027) was coated on it.
For patterning electrodes, we used Electron Beam Lithography Exposure. The electrodes were completed bydeveloping and etching. The flow channel was made of PDMS resin and laminated the substrate and PDMS resin
to fabricate the EHD pumps.

Experimental Method
Figure 2 depicts a schematic of the experimental apparatus. EHD pump, a differential pressure gage (Validyne
Engineering, DP-15-22-N-1-W4-A), and a flow mater (Sensirion, SLI2000) were used in the experiment. For eachcomponent, 1/4 inch tubes were used for the connection. A differential pressure gauge connected to the pump’s
upstream and downstream was used to measure the pressure generated by the pump. By closing the valve, wewere able to conduct experiments with no net flow. The current was measured by the resistor (320 kΩ) inserted
between the EHD pump and the power supply. The resistance value of our EHD pump (L2 = 50 μm) was about 1
MΩ. During the experiments, the pressure and electrical current were monitored at 1-second intervals by a datalogger. Stepwise, DC voltage was applied to the electrodes of the EHD pump in the experiment. The maximum
apparent electric field strength between the electrodes, defined as the applied voltage divided by the electrode gap,
was 6 kV/mm.

NUMERICAL MODEL
Governing Equations for EHD
We used COMSOL Multiphysics for conduct numerical simulation. The flow was assumed to be two-dimensional
at steady-state and incompressible in this study, and the flow field was analyzed using the Navier-Stokes equation
shown as Eq. (1) as well as the continuity equation of Eq. (2).
𝜇𝑓
1(𝐮 ∙ ∇)𝐮 = − ∇𝑃 + ∇2 𝐮 + 𝐅
𝜌
𝜌

(1)

∇∙𝐮=0

(2)

Where u is the velocity vector, P is the pressure, ρ is the fluid density, μf is the viscosity coefficient, and F is the
body force vector. The equation of electric field as shown below must also be solved to calculate the volume force
of Coulomb force. Eq. (3) and Eq. (4) are the conservation laws of positive and negative dissociated charges,
respectively.
𝛻 ⋅ (𝑞𝜇 + 𝐄 + 𝑞𝐮 − 𝐷𝑖 𝛻𝑞) = 𝑘𝑑 𝑐 − 𝑘𝑟 𝑞𝑤 = 𝑘𝑟 (𝑤0 2 𝐹(𝜔) − 𝑞𝑤)(3)𝛻 ⋅ (−𝑤𝜇 − 𝐄 + 𝑤𝐮 − 𝐷𝑖 𝛻𝑤) = 𝑘𝑑 𝑐 − 𝑘𝑟 𝑞𝑤 = 𝑘𝑟 (𝑤0 2 𝐹(𝜔) − 𝑞𝑤)(4)

3

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Figure 2. Schematic of experimental apparatus.

Eq. (5) shows Gauss’s law. Eq. (6) shows relationship between the electric field and electric potential. Eq. (7)shows Coulomb force. Eq. (8) shows charge density at equilibrium. Eq. (9) and Eq. (10) show the Onsager function
representing field-enhanced dissociation. Eq. (11) shows the rate of recombination.
𝛻∙𝐄= −𝑞−𝑤
𝜀(5)

𝐄 = −∇𝜙

(6)

𝐅=

𝑞
𝐄
𝜌

(7)

𝑤0 =

𝜎
2𝜇

(8)

𝐹(𝜔) =𝐼1 (4𝜔)2𝜔

(9)

𝜔(|𝐄 ∗ |) = 𝛽𝐶0 −1⁄2 |𝐄 ∗ |−1⁄2
𝑘𝑟 =

2𝜇
𝜀

(10)
(11)

In these equations, μ₊ and μ₋ are the ionic mobilities of the positive and negative charges, respectively. Assuming
they are equal, we use μ for the ionic mobility. Di is the ion diffusion coefficient, E is the electric field strength,
ϕ is the electric potential, q is a positive charge density, w is a negative charge density, c is the neutral molecule
concentration, kr is the recombination coefficient, kd is the dissociation constant, σ is the conductivity, ε is therelative permittivity, and I1 is the Bessel function of the first kind.

Nondimensionalization of The Governing Equations
The nondimensionalized governing equations derived by using nondimensional parameters are shown in Eqs.
(12) to (16). The nondimensional parameters are shown in Eqs. (17) to (21). Eq. (22) represents the dissociation
tendency.
𝜕𝐮∗1+ (𝐮∗ ∙ ∇)𝐮∗ = −∇𝑃 ∗ + 𝐸 ∇2 𝐮∗ + 𝑀𝟐 𝐶𝟎 (𝑞 ∗ − 𝑤 ∗ )𝐄 ∗
∗
𝜕𝑡
𝑅𝑒

(12)

∇ ∙ 𝐮∗ = 𝟎

(13)

4

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The 11th JFPS International Symposium on Fluid Power 2020∇∗ ∙ (𝑞 ∗ 𝐄 ∗ + 𝑞 ∗ 𝐮∗ − 𝛼∇∗ 𝑞 ∗ ) =𝐶0 2 𝐹(𝜔)1− (1 + ) 𝑞 ∗ 𝑤 ∗1
𝑏𝑟(1 + )𝑏𝑟(14)

𝐶0 2 𝐹(𝜔)1− (1 + ) 𝑞 ∗ 𝑤 ∗1
𝑏𝑟(1 + )
𝑏𝑟(15)∇∗ ∙ (−𝑤 ∗ 𝐄 ∗ + 𝑤 ∗ 𝐮∗ − 𝛼∇∗ 𝑤 ∗ ) =𝐶0 =∇2 𝜙 ∗ = −𝐶0 (𝑞 ∗ − 𝑤 ∗ )

(16)

𝜎0 𝐿2
𝐿2
1 𝐿2
=
=
𝜀
2𝜀𝜇𝐸 2𝜇𝐸 ( ) 2 𝐿𝑡𝑟𝑎𝑣𝑒𝑙
0 𝜎

(17)

𝑘𝐵 𝑇
𝑒0 𝐸0 𝐿2

(18)

𝑅𝑒 𝐸 =

𝜌𝜇𝐸0 𝐿2
𝜇𝑓

(19)

𝑀=

√𝜀 ⁄𝜌
𝜇

(20)

𝜇+
𝜇−

(21)

𝛼=

𝑏𝑟 =

𝛽=(

1⁄2

𝑒0 3 𝜎𝐿2
32𝜋𝜀 2 𝜇𝑘𝐵 2 𝑇

(22)

)
2

In these equations, E0 is strength of electric field, e0 is elementary charge, kB is Boltzmann constant, and T is thetemperature. Conduction number, C0 is the ratio of the length between the electrodes to the length that the ions
can travel by migration within the relaxation time which it takes for the dissociated charges to recombine. α is the
diffusion number. ReE is electric Reynolds number. It includes the ionic drift velocity μE0 as the representative
velocity and is not hydrodynamic Reynolds number. M is mobility number. It is the ratio of hydrodynamic
mobility to the ionic mobility. It depends on only liquid properties. These numbers were introduced in a previous
research [20]. br is mobility ratio of the positive and negative dissociation charges. β is the parameter that accounts
for the electric field dissociation of dissociative molecules. The larger β, the more likely it is that dissociated ions
are generated. Nondimensional parameters and liquid properties that we used are shown in Table 2. These
properties were obtained from [21].
Table 2. Nondimensional quantities and liquid properties for numerical simulation
Nondimensional numbers
L2 = 500 µm
L2 = 50 µm
L2 = 5 µm
Conduction number C0 (-)

0.50 - 1.50

0.050 - 0.15

0.005 - 0.015

Diffusion number α (-)

0.08 × 10-7 - 0.25 × 10-7

0.08 × 10-6 - 0.25 × 10-6

0.08 × 10-5 - 0.25 × 10-5

Electric Reynolds number ReE (-)

1.15×102 - 3.47 ×102

1.15 ×10 - 3.47×10

1.15 - 3.47

0.50

0.16

0.05

Nondimensional number β (-)
Liquid properties (Vertrel XF)
Density ρ (kg/m3)

1580

Conductivity σ (S/m)

3.45×10-8

Relative permittivity ε/ε0 (-)

6.84

Viscosity μf (Pa‧s)

6.47×10-4

Ionic mobility μ (m2/Vs)

4.64×10-8

Ion diffusion coefficient Di (m2/s)

1.2×10-9

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RESULTS AND DISCUSSION600ΔPΔP (Pa), Current (μA)

500
400

I

300
V

200
100
0
0

500

1000
900
800
700
600
500
400
300
200
100
0
2000

Voltage (V)The time variations of the pressure rise, ∆P, and electric current, I, of the pump with L2 = 50 µm and h = 4L2 areshown in Figure 3. The experiment was stopped at 300 V to avoid breakdown of the fluid. The power consumption
was 0.1 W at 300 V.
As shown in Figure 4, the pressure rises of all pumps increased with the electric field. Due to differential pressure
gage overpressure, the pressure rises of the pump with L2 = 50 µm were not measured at 6 kV/mm. The maximum
generated pressure averaged 727 Pa at the channel height of 50 µm, 463 Pa at 200 µm, and 20 Pa at 1000 µm asshown in Figure 4. The results showed that a higher pressure rise was generated by a lower channel height. The
reason for this characteristic is that charges are distributed on a large percentage of cross-section of the channel
with a lower height as shown in Figure 5.
Figure 6 depicts the characteristic of the ∆P vs Electric field strength. Figure 6 show that the pump with L2 = 50µm generated higher pressure than the other pumps. The pump with L2 = 5 µm is exceedingly small. The
manufacturing accuracy of the pump was not accurate due to its exceedingly small scale, and which may haveaffected the experimental results. In Figure 7, the effects of nondimensional parameters C0 and β on the generatedpressure are shown including the results of the previous research [20], in which the present results are circled (ourresult of β = 0.05 is not plotted because generated pressure was very small and parallel plate electrode was usedin previous research [20] ). Figure 6 and Figure 7 shows that our EHD pumps could generate higher pressure thanthe previous research and generated pressure increase as the electric field strength increase. However, the resultsof previous research show that there are areas where generated pressure is not proportional to the electric fieldstrength under small C0 and β conditions. Therefore, we will build a smaller EHD pump than this research andconduct experiments to investigate pump characteristics.

1000
1500
Time (s)Figure 3. Variations of pressure and electric current of the pump with L2 = 50 µm and h = 200 µm at zero flow rate.

1000

Series3
50 μm
Series2
200 μm1000 μm
Series1

ΔP (Pa )

800
600
400
200
0

0

1
2
3
4
5
6
7
Electric field strength ( kV/mm )Figure 4. Effect of channel height on pressure rise at zero flow rate of the pump with L2 = 50 µm.

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The 11th JFPS International Symposium on Fluid Power 2020(a)h = 50 µm.

(b) h = 200 µm.

ΔP (Pa )(c) h = 1000 µm.
Figure 5. Charge density distribution at zero flow rate. The unit is C/m3.

500
450
400
350
300
250
200
150
100
50
0

L2 = 500 μm
L2 = 50 μm
L2 = 5 μm

0

1

2

3

4

5

6

7

Electric field strength ( kV/mm )Figure 6. Characteristic of ∆P vs Electric field strength.

7

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10003
10

10100210

ΔP (Pa)

101
1001
0.1

10-1

100.01-2
0.001
-3
10

-4
0.000110
1.00E-03
10-3

10-2

1.00E-02

10-1

1.00E-01

0
1.00E+00

10

C0

101

1.00E+01

2
10
1.00E+02

103

1.00E+03

Figure 7. Comparison of C0 and β between the present research and previous one (present results are circled).

Numerical simulation results are shown in Figures 8 to describe the difference for each similar scale pump. Figure
8 (a) to (c) depict the charge density distribution, while (d) to (e) depict the velocity distribution. In Figure 8 (a)to (c), red and blue lines show the outer lines of each heterocharge layer which is defined as the area where thecharge density is higher than equilibrium one, w0. Here, the overlapping of heterocharge layers and smaller-scale
EHD pump have larger area of the overlap. This is because the travel distance of dissociated ions within therelaxation time is independent of electrode scale as shown in Eq. (17). This behavior of the charge distribution
on each scale is thought to be related for the difference of the pump characteristic. Also, electric charges ofsmaller-scale pump distributed on a large percentage of cross-section of the channel. Especially, for the smallest
scale pump, the heterocharge layer is over the boundary of the computational domain. Figure 8 (d) to (f) showthat the velocity distribution was differed significantly depending on the pump scale. In the pumps with L2 = 500µm and L2 = 5 µm, the vortices were generated near the electrodes only. However, in the pump with L2 = 50 µm,
the vortices were generated at various locations of channel.

(a) Pump with Pump with L2 = 500 µm, h = 2000 µm.

(d) Pump with Pump with L2 = 500 µm, h = 2000 µm.

(b) Pump with Pump with L2 = 50 µm, h = 200 µm.

(e) Pump with Pump with L2 = 50 µm, h = 200 µm.

(c) Pump with Pump with L2 = 5 µm, h = 20 µm.

(f)

Pump with Pump with L2 = 5 µm, h = 20 µm.

Figure 8. (a)~(c) Electric charge density at zero flow rate. The unit is C/m3.
(d)~(f) Velocity distribution at zero flow rate. The unit is m/s.

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CONCLUSION
In this research, characteristics of microscale EHD pumps with different channel heights and sizes were
investigated. Generated pressure depends on channel height and we could obtain higher pressure by lowering thechannel height. The flow rate for lower channel height should be measured because lower channel height may
decrease flow rate. By numerical simulation, overlap of the positive and negative heterocharge layers wasobserved at L2 = 50 µm or less. Therefore, it is necessary to design the electrode considering the overlap of theheterocharge layer on microscale EHD pumps.

ACKNOWLEDGMENTS
This research was partially supported by Toukai Foundation for Technology, the Yashima Environment
Technology Foundation, and the JSPS KAKENHI Grant Number 21K14075. The authors appreciate Mr. T.
Hizawa of Toyohashi University of Technology’s Evolution of the Electronics-Inspired Interdisciplinary ResearchInstitute (EIIRIS) for his assistance with MEMS fabrication.

NOMENCLATURE
br

Mobility ratio of the positive and negative dissociation charges (-)

c

Neutral molecule concentration (mol/m3)

C0

Conduction number (-)

Di

Ion diffusion coefficient (-)

e0Elementary charge (C)E

Electric field strength vector (V/m)

E0Electric field strength (V/m)

F

Body force vector (N/kg)

F(ω)

Onsager function (-)

I1

Bessel function of the first kind (-)

kB

Boltzmann constant (J/K)

kd

Dissociation constant (C/mol‧s)kr

Recombination coefficient (m3/C‧s)

L2

Electrode gap (m)LtravelThe distance of that ion travel (m)M

Mobility number (-)∆P

Generated pressure (Pa)q

Positive charge density (C/m3)

ReE

Electric Reynolds number (-)

T

Temperature (K)uVelocity (m/s)

uVelocity vector (m/s)

w

Negative charge density (C/m3)

9

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w0

Charge density at equilibrium (C/m3)

Greek
ε

Permittivity (F/m)μ

Ionic mobility (m2/Vs)

μf

Viscosity (Pa‧s)σ

Conductivity (S/m)ϕ

Electric potential (V)αDiffusion number (-)β

Nondimensional number with respect to the Onsager function (-)ρ

Fluids density (kg/m3)

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The 11th JFPS International Symposium on Fluid Power 2020

Study on a Low Pressure and Flowrate Driving ofMicro Leg Joints for Soft Robots
Koji MICHISHITA*, Kazuhiro YOSHIDA**, and Joon-wan KIM***

School of Engineering, Tokyo Institute of Technology**
FIRST, IIR, Tokyo Institute of Technology
J3-12, 4259 Nagatsuta-cho, Midori-ku, Yokohama, 226-8503 JAPAN(E-mail: woodjoon@pi.titech.ac.jp)

Abstract. To aim at developing a micro leg joint driven by low pressure and flowrate, we propose a novel soft
actuator composed of extremely soft silicon rubber (Ecoflex) and reinforced by a fiber. Due to two different
flexibilities of rubber materials (Ecoflex: stretchable material; PDMS: structural material), the proposed balloon-type
actuator generates angular displacement like an insect leg joint at the low applied driving pressure. This actuator isalso designed to be driven at a lower flowrate by attaching a fiber to the balloon and restricting its expansion in
directions unrelated to the angular displacement. The actuator was successfully fabricated by the molding and curingof rubbers with the attachment of the fiber. To verify the influence of the fiber, we measured the actuation angle andits volume change. The proposed fiber-reinforced actuator achieved the same actuation angle with less volume
change compared with one of non-fiber actuators. This study proved the feasibility that the reduction of volume
change can increase the dynamic response of the soft actuator.
Keywords: Electro-conjugate fluid (ECF), Soft robotics, Micropump, MEMS, Functional fluid

INTRODUCTION
In recent years, the use of microrobots is expected to search for minute spaces that humans cannot enter, such as
under the rubble of disaster sites. It is known that a robot that walks with multiple legs rather than wheels orcaterpillars is suitable for moving on rough terrain because of its high adaptability to the ground [1]. The
walking of a multi-legged robot composed of rigid materials requires a large number of sensors and complicatedcontrols. On the other hand, soft robots made of flexible materials have the advantage of conformity to objects
due to their deformation and the absorption of impact energy, so that they can walk on rough terrain without theneed for complicated control. To integrate our ECF micropumps [2] as hydraulic power sources into soft robots,
we need to develop their actuators to be driven with low pressure as well as low flow rate.
This study proposes a fiber-reinforced soft microactuator that generates angular displacement like the leg jointof insects with the low pressure and low flow rate driving. The structure and fabrication process of the actuator
are described in this study. Also, its characteristics are clarified by the simulations and experiments.

FIBER-REINFORCED SOFT JOINT ACTUATOR
As a way to reduce the driving pressure, we propose to fabricate the soft actuator as a micro leg joint by
utilizing two kinds of rubber with different elastic modulus: Ecoflex as a stretchable material and PDMS as astructural material, as shown in Fig. 1(a). This multi-rubber structure is adequate for low pressure driving, but
inadequate for low flowrate driving, because a large amount of driving fluid does not contribute to thedisplacement, as shown in Fig. 1(b).
To satisfy the driving with low flowrate as well as low pressure, this study proposes to limit the deformation that
does not contribute to the operating angle by using reinforcing fibers. Fig. 2 shows the structure and working
principle of the fiber-reinforced soft joint actuator. The actuator has a hollow structure composed of hard rubber(PDMS), soft rubber (Ecoflex), and reinforcing fibers wrapped around the soft rubber. When the actuator ispressurized, the soft rubber deforms more than the hard rubber. As a result, the hard rubber becomes adisplacement constraint and creates an angular displacement. A reinforcing fiber plays a role in suppressing only
the deformation of the soft rubber in the vertical direction of the wall surface as shown in Fig. 2(a), reducing thevolume change of the hollow portion during the actuator operation. This fiber reinforcement can reduce theamount of working fluid to enter the hollow portion, improving the dynamic characteristics.
Fig. 3 shows the fabrication and assembling process of the fiber-reinforced soft joint actuator. For hard rubber,
PDMS (Sylgard 184, Toray Dow Corning) and its curing agent are mixed at a mass fraction of 10: 1. For soft
rubber, Ecoflex (Ecoflex00-30, Smooth-On) is used. The lower part and the ceiling are fabricated separately.

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The 11th JFPS International Symposium on Fluid Power 2020

Siliconetube
0.5 mmA

5.2 mmA’A

A’
Hard rubber（PDMS）

0.5 mm

0.4 mmSoft rubber(Ecoflex)(a) Soft joint actuator composed of two rubber materials(b) Problem of the micro leg joint

Figure 1. Structure, working principle, and problem of the soft joint actuator made of Ecoflex and PDMS

Reinforcing fiber(a) Structure and working principle(b) Limiting the deformation unrelated to the angular displacement

Figure 2. Proposed fiber-reinforced soft joint actuator(a) Forming PDMS
sidewalls(b) Forming Ecoflex
sidewall(e) Assembling(c) Forming Ceiling(f)

Detachmentfrom molds(g) Attachment ofreinforcing fiber(d) Cutting tube

Figure 3. Fabrication and assembling process of the fiber-reinforced soft joint actuator
Fig.2 Fabrication procedure of the proposed fiber-reinforced soft actuator

After bonding the ceiling and the lower part of the actuator, both ends of the reinforcing fiber are fixed to theceiling of the actuator with a silicone adhesive. There must be no silicone adhesive between the fiber and thesidewall of the actuator not to increase the thickness of the sidewall.

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SIMULATIONS AND EXPERIMENTS
Fig. 4 shows the operation of the soft joint actuators at the applied pressure of 30 kPa with a wall thickness of
0.4 mm. Similar to the simulation results, it was confirmed the angular displacement was caused by thedifference in elastic modulus between PDMS and Ecoflex. In addition, it can be seen that the Ecoflex part of the
non-reinforced soft joint actuator expands in the vertical direction of the wall surface, as in the simulation. On
the other hand, in the fiber-reinforced soft joint actuator, it was confirmed that the expansion of the Ecoflex partwas suppressed by the reinforcing fiber. Moreover, the deformation form was in good agreement with thesimulation by COMSOL MultiphysicsTM. Fig. 5(a) shows the simulation and experiment results between theapplied pressure and the actuation angle under each fabrication condition. As for the tendency of the change ofthe actuation angle with respect to the increase of pressure, the experimental value and the simulation result
were almost the same. In addition, the angles at the applied pressure of 30 kPa with the wall thickness of 0.4
mm were 24.2 ° and 23.6 ° without and with the fiber reinforcement, respectively. Their actuation angles were
almost the same as each other. Fig.5(b) shows the relation between the applied pressure and the volume changeof the actuator. Comparing the volume change of the actuator without and with the fiber reinforcement, thevolume change of the fiber-reinforced actuator became significantly smaller. The fiber reinforcement makes it
possible to drive the actuator with a smaller amount of fluid, improving its dynamic characteristics.

von Mises stress [Pa]30 kPa

30 kPavon Mises stress [Pa]30 kPavon Mises stress [Pa]30 kPaReinforcing fiber(a) Without the fiber reinforcement

von Mises stress [Pa]30 kPa5 mm(b) With the fiber reinforcement

Figure 4. Comparison of the soft joint actuator without and with a fiber reinforcement

Simulation― t = 0.3 mm― t = 0.4 mm― t = 0.5 mm

010Experiment● t = 0.3 mm● t = 0.4 mm● t = 0.5 mm

20

30

40

Internal pressure [kPa]

50

50
45
40
35
30
25
20
15
10
5
0

Volume change [mm3]Non-reinforced

Actuation angle [ ˚ ]Actuation angle [ ˚ ]

140
50
45
40
35
30
25
20
15
10
5
0

Fiber-reinforced
Simulation― t = 0.3 mm― t = 0.4 mm― t = 0.5 mm

Experiment▲ t = 0.3 mm▲ t = 0.4 mm▲ t = 0.5 mm

Non-reinforced● t = 0.3 mm● t = 0.4 mm● t = 0.5 mm

120
100

Fiber-reinforced▲ t = 0.3 mm▲ t = 0.4 mm▲ t = 0.5 mm

80
60
40
20
0

0

10

20

30

40

Internal pressure [kPa](a) Actuation angle without and with the fiber reinforcement

50

0

10

20

30

40

50

Internal pressure [kPa](b) Necessary volume change

Figure 5. Actuation angles and necessary volume changes of the soft joint actuators

CONCLUSIONSIn this study, we proposed the structure and fabrication process of a fiber-reinforced soft joint actuator thatgenerates angular displacement like the leg joint of an insect with low pressure and flowrate driving. The
proposed fiber-reinforced soft joint actuator was successfully fabricated and the fiber-reinforcement succeededin suppressing unnecessary expansion while obtaining the same actuation angle as non-reinforcement.

REFERENCES
1. MAHAPATRA, A., ROY, S. S., PRATIHAR, D. K., Multi-body Dynamic Modeling of Multi-legged Robots, Springer,
Singapore, 2020, pp.1-8.
2. KIM, J., TANABE, Y., YOKOTA, S., Comprehending Electro-conjugate Fluid (ECF) Jets by Using the Onsager Effect,
Sensors and Actuators A: Physical, 2019, 295-15, pp.266-273.

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Organized Session | Organized session

Simulation and Modeling[OS3-01] A Study on the Virtual Simulation Model of an Excavator Equipped with a
Tiltrotator Based on Simscape〇Seongwoong Choi1, Kyungsin Kwak1, Yongseok Kim1, Kyoungkwan Ahn1, Soonyong Yang1（1.University of Ulsan）[OS3-02] Modeling, Simulation, and Control of Blade Pitch to Improve the
Performance of a Hydrostatic Wind Turbine〇Neil Christopher Garcia1, Kim Stelson1 （1.University of Minnesota）[OS3-03] Internal Flow and Hysteresis Characteristic of the Poppet Type Pressure
Control Valve〇Seiei Masuda1, Fumio Shimizu2, Masaki Fuchiwaki2, Kazuhiro Tanaka2 （1.Control Systems
Engineering Department, IHI Corporation, 2.Graduate School of Computer Science and Systems
Engineering, Kyushu Institute of Technology）[OS3-04] Flow Patterns and Hysteresis Characteristic of a Poppet Valve〇Naoki Hirose1, Seiei Masuda2, Fumio Shimizu3, Masaki Fuchiwaki3, Kazuhiro Tanaka3 （1.Voith
IHI Paper Technology, 2.IHI Corporation, 3.Kyushu Institute of Technology）[OS3-05] Numerical Study on Identification Input for Nonlinear Hydraulic Arms.
〇Teruo Kato1, Satoru Sakai1, Ryo Arai1 （1.University of Shinshu）[OS3-06] On the Analysis of Energy Behaviors in Hydraulic Cylinder Dynamics via
Modeling of Experimental Excavators〇Ryo Arai1, Satoru Sakai1, Akihiro Tatsuoka2 （1.Shinshu University, 2.Mitsubishi Heavy
Industries）©The Japan Fluid Power System Society

OS3-01

The 11th JFPS International Symposium on Fluid Power 2020

A Study on the Virtual Simulation Model of an Excavator Equippedwith a Tiltrotator Based on SimscapeSeongwoong CHOI*, Kyungsin KWAK*, Yongseok KIM**, Kyoungkwan AHN**,
Soonyong YANG***

Department of Construction Machinery Engineering, General Graduate School, University of Ulsan
93, Daehak-ro, Nam-gu, Ulsan, Republic of Korea(E-mail: soonyy@ulsan.ac.kr)**
Department of Mechanical Engineering, School of Mechanical Engineering, University of Ulsan
93, Daehak-ro, Nam-gu, Ulsan, Republic of Korea

Abstract. In this paper, we develop a simscape-based simulation model and modeling of excavator with tiltrotator.
Dynamic model of excavator 3D models using Simscape Multibody and Hydraulic model of excavator using
Simscape Hydraulics were modeled. We apply PID controller to make control system that works along the input
angle. Simulation results show that errors occur when comparing input and output angles, but overall, they movealong input angles.
Keywords: Construction Machinery, Excavator, Tiltrotator, Simscape, Simulation.

INTRODUCTION
The excavator is a representative equipment working in the construction site. However, it is also used in various
fields such as agriculture and forestry, and recently, the scope of use is expanding to the submarine area.[1][2]
Typical operations of an excavator include excavation work and leveling work. Excavation work is performed ina variety of workspaces, such as changing field usage, changing drains and pipes, and removing buried soil. The
planarization work is performed in a large space such as solidifying the foundation of the field and leveling theequipment access road. This work has a limitation on the working posture of the existing excavator. Therefore,
by installing a tiltrotator serving as a wrist, unnecessary operations can be reduced to perform work efficiently
and work time can be reduced. Due to this advantage, as the number of excavators equipped with a tiltrotator
increases, various studies related to this have been conducted, and research and development on automation,
autonomy, and electricization have recently become an issue in relation to excavators.[3] Therefore, in thispaper, a study on a virtual simulation model of an excavator equipped with a tiltrotator for automation andautonomy based on Simscape was conducted.

SIMULATION MODEL OF VIRTUAL EXCAVAOR
The simulation model operates the hydraulic valve and actuator of the model by inputting the input angle asshown in figure 1 and receiving the valve control signal through the controller. The excavator model operates bytransmitting the force and torque generated by the actuator to the multibody model.

Figure 1. Configuration of Excavator Simulation Model

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The 11th JFPS International Symposium on Fluid Power 2020

Excavator Model
To model the hydraulic model, which is the power source of the excavator, and the dynamic model representing
the force and torque applied to each joint, it was modeled using Simscape Multibody and Simscape Hydraulics
libraries.

3D ModelAs shown in Figure 2, the upper body (cabin) and lower body (crawler), boom, arm, bucket, tilt rotator, andcylinder of the excavator were modeled using CATIA.

Figure 2. 3D model of excavator with tiltrotator

Multibody Model
The dynamic model representing the excavator's movement was implemented with the force and torque of eachjoint using Simscape Multibody. The 3D shape of the excavator was modeled using CATIA and Solidworks. To
convert to a Simscape Multibody model, create a Simscape Multibody model using the XML file conversion
function in Solidworks and create a Data file of the Assembly. The modeled multibody model is shown in
Figure 3.[4][5]Figure 3. Multibody Model of Excavator

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Hydraulics Model
The hydraulic model of the excavator, the main pump, MCV, and actuator (cylinder, motor) were modeled using
Simscape Hydraulics. The modeled hydraulic model is shown in Figure 4.[5]Figure 4. Hydraulics Model of Excavator

Controller & Final Model
It was composed of a feedback controller using the error of each joint input angle and output angle of the
excavator. Among the feedback controllers, it was composed of a PID controller, and it was operated using anangle error. An appropriate value was selected by adjusting the gain value of each joint. The controller model isshown in Figure 5. The simulation model was finally completed by combining the previously constructedexcavator model and the controller model. The final simulation model is shown in Figure 5.

Figure 5. Controller and Final Simulation Model of Excavator

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SIMULATION AND RESULTS
The simulation of the completed simulation model proceeded as follows. A simulation was performed for the
single motion of the excavator boom, arm, bucket, and tiltrotator, and the input angle of each joint was set asshown in Figure 6. The simulation result for this is shown in Fig. 7, and in the result graph of Fig. 7, the blueline is the input angle and the red line is the output angle. In the resulting graph, you can see that it moves along
the input angle over the entire range of motion limit angles, except for differences that occur in some parts. It is
expected that the error occurring here will be solved by applying gain adjustment and a control algorithm thatcomplements it.

Figure 6. Simulation Input Angle of Excavator Single Operation

Figure 7. Simulation Results of Excavator Single Operation

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CONCLUSION
In this study, a virtual simulation model of an excavator was developed using Simscape of MATLAB. The
developed model confirmed the posture and trajectory of the excavator operated by the input angle. The
operation of the simulation model was followed according to the input angle, but some slight errors appeared.
The error occurring here is thought to be caused by the speed change occurring in the valve direction change. It
seems that the error can be solved by applying a control algorithm that can compensate for the gain adjustmentof the controller. In future research, we will proceed with research on the control algorithm to improve thedeveloped simulation model.

ACKNOWLEDGMENTS
This paper was supported by Korea Institute for Advancement of Technology(KIAT) grant funded by the Korea
Government(MOTIE) (P0012769, The Competency Development Program for Industry Specialist)

REFERENCES
1. T. H. LIM, H. S. LEE and S. Y. YANG, Development of Simulator of Hydraulic Excavator, Journal of Drive andControl, 2005, 2-1, pp.63-68.
2. T. H. LIM and S. Y. YANG, Development and Application of Simulator for Hydraulic Excavator, Journal of the Korean
Society for Precision Engineering, 2006, 23-9, pp.142-148.
3. S. W. CHOI, Q. H. LE, T. G. SON and S. Y. YANG, A Study on Construction of Control System for Wireless Remote
Control of Small Excavator, Journal of Drive and Control, 2020, 17-4, pp.103-112.
4. D. S. AHN, I. Y. LEE and H. H. KIM, Integrated SolidWorks & Simscape Platform for the Model-Based Control
Algorithms of Hydraulic Manipulators, Journal of Drive and Control, 2015, 12-4, pp.41-47.
5. Q. H. LE, Y. M. JEONG, C. T. NGUYEN and S. Y. YANG, Development of a Virtual Excavator using SimMechanics
and SimHydraulic, Journal of Drive and Control, 2013, 10-1, pp.29-36.

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Modeling, Simulation, and Control of Blade Pitch to Improve the
Performance of a Hydrostatic Wind Turbine
Neil GARCIA and Kim A. STELSON
Department of Mechanical Engineering, College of Science and Engineering, University of Minnesota,
111 Church St SE, Minneapolis, MN 55455, USA(E-mail: garc0408@umn.edu, kstelson@umn.edu)
Abstract. Hydrostatic transmissions are a valuable addition to community-scale wind turbines, reducing maintenancedemands and allowing for compact energy storage with comparable efficiency to gearbox transmission units. In thispaper we create a model to simulate the performance and improve the control of blade pitching for a wind turbine
fitted with a hydrostatic transmission. A model of a 60-kW hydrostatic wind turbine with K⍵2 torque control waspreviously created in MATLAB Simulink. To add blade pitching into the existing model, a wind turbine blade wasdigitally recreated, simulated, and used to fit a general equation for the turbine’s power coefficient. The model was
validated by comparing results with the previous model. A modified proportional controller for blade pitch actuation
was added to the system to capture extra energy and dissipate excess energy.
Keywords: hydrostatic transmission, hydrostatic wind turbine, turbine blade pitch

INTRODUCTION
As demand for clean energy alternatives rises around the world, improvements to current strategies for harnessing
wind energy have become increasingly valuable. Modern utility-scale wind turbines are equipped with sensors
and actuators to adapt to dynamic wind conditions, optimizing power generation and maintaining safe operating
conditions with increasingly advanced control schemes.
More sophisticated transmission systems make wind power more reliable and economical. Most wind turbines
use transmissions to convert high torque, low speed rotor rotation into high speed, low torque generator rotation.
Unlike traditional wind turbine transmissions, which rely on heavy fixed-ratio gearboxes and power electronics[8], a hydrostatic transmission (HST) uses hydraulic lines to transmit power where the rotation of the turbine rotor
drives a pump in circuit with a variable displacement hydrostatic motor as shown in Figure 1. The motor thendrives a synchronous generator, adjusting its swash plate angle to match the electric grid frequency [6].
A conventional gearbox transmission with a doubly fed induction generator has 95% transmission efficiency, a
96% generator efficiency, and a 97% power electronics efficiency, for a total efficiency of 88.5%. A hydrostatictransmission has an 87% transmission efficiency and a 98% generator efficiency for a total efficiency of 85.2%[10]. However, the continuous variability of a hydrostatic transmission, the elimination of power electronics, thereliability of the components and the improved power-to-weight ratio are all attractive merits of HSTs in windturbines.

Figure 1. Hydrostatic wind turbine transmission [6].

The Hydraulic Power Transmission Lab at the University of Minnesota has created a simulation of a 60-kW windturbine (the Polaris P21-60) outfitted with the lab’s hybridized hydrostatic transmission, using wind and rotor
speed inputs to calculate rotor torque. Current studies in the lab are investigating the use of a hydraulic1

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accumulator to store excess energy otherwise wasted [6] and advanced control strategies such as extremum
seeking control to maximize power generation.
While the effect of blade pitching on power generation in conventional wind turbines is well understood, theimpact of adding blade pitching to the hydrostatic wind turbine control with hybridization is not fully understood.
The goal of this research is to add blade pitching to the existing simulation as a tool to make system performance
predictions and improve the hydrostatic wind turbine control strategy.

SYSTEM MODELINGHYDROSTATIC TRANSMISSIONS IN WIND TURBINES
The hydrostatic transmission of a wind turbine is governed by the dynamics of the rotor shaft and HST linepressure as described by Eqs. (1) and (2) [6]:𝐽𝑟 𝜔̇ 𝑟 = 𝜏𝑟𝑜𝑡 (𝜔𝑟 , 𝑢, 𝛽) − 𝑏𝑟 𝜔𝑟 − 𝐷𝑝1 𝑝ℎ𝑠𝑡
𝑝̇ℎ𝑠𝑡 =𝛽𝑒𝑉

(𝐷𝑝1 𝜔𝑟 − 𝐿𝑚 𝑝ℎ𝑠𝑡 − 𝛼𝐷𝑚 𝜔𝑔 )

(1)
(2)

𝜏rot is aerodynamic torque, a function of rotor speed ⍵r, wind speed u, and blade pitch 𝛽. br is the combined
damping of the rotor shaft and pump. Dp1 is the displacement of HST pump. phst is the line pressure of the HST.
𝛽e is the effective bulk modulus of the fluid and line, linearized at operating line pressure. V is the fluid line
volume. Lm is the sum of losses in the main pump and motor, assuming losses in the line are negligible. Dm is thedisplacement of the motor. α is the normalized swash plate angle of the motor. ⍵g is the generator shaft speed.

FUNDAMENTALS OF WIND TURBINE CONTROLWind turbines operate in four regions dependent on wind speed and turbine characteristics [8], shown in Figure
2. In Region 1, wind speed is too low for turbine operation, and blades are pitched to minimize aerodynamic
torque. In Region 2, power is generated, but operation is below the rated wind speed. Rotor torque control is theprimary mode of power regulation, but turbine blades are also pitched to an angle 𝛽 to maximize power generation.
The power coefficient of the turbine Cp is a function of 𝛽 and the tip speed ratio (TSR) λ:
𝜆=𝑅

𝜔𝑟

(3)

𝑢

where R is turbine blade radius. To maximize power in Region 2, rotor torque is adjusted according to the K⍵2control law [7], where the turbine operates with torque 𝜏c to maximize energy capture. Desired torque is given by𝜏𝑐 = 𝐾𝜔𝑟 2
1

𝐶

𝐾 = 𝜌𝜋𝑅5 𝑝,𝑚𝑎𝑥
3
2

𝜆𝑠

(4)
(5)

where 𝜌 is the density of air, Cp, max is the maximum power coefficient of the wind turbine, and λs is the ideal TSR.
In Region 3, operation is above rated wind speed and turbine control seeks to maintain safe operating conditions
by keeping power constant, pitching the blades out of the wind to dissipate excess energy and keep power
generation at rated power. In Region 4, far above rated wind speed, wind speeds are too high for operation, so
blades are pitched completely out of wind and the rotor is braked to prevent system damage.

Figure 2. Wind turbine power curve, labeled with regions of operation.

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The existing MATLAB Simulink model of the hydrostatic wind turbine includes pressure and rotor dynamics
and K⍵2 control (adjusting swash plate angle α to optimize torque for power). Wind inputs and rotor speed
are translated to rotor torque by referencing a 2-D interpolation table for 𝜏rot (u, ⍵r) shown in Figure 3. A
constant blade pitch of 0° is assumed as optimal in Region 2, although additional energy can be captured byusing non-zero pitch. In Region 3, blades are pitched to limit power, with the actuation assumed to beinstantaneous.

Figure 3. Visualization of the interpolation table for torque for the simulated 60-kW turbine [5].

BLADE PITCH SIMULATION
To accurately include blade pitch in the model, the existing interpolation table needed refinement. To do so, theturbine blades were modeled and simulated in QBlade. QBlade [4] is a blade element momentum analysis programthat allows for the construction, visualization, and simulation of wind turbine blades. QBlade uses multi-parameterblade element analysis to simulate blades for a range of wind speeds, rotational speeds, and blade pitches. A
general equation incorporating blade pitch can be fit to simulation results and used in dynamic simulation andcontrol models. Ideally, the blade designed and simulated in QBlade would directly emulate the 60-kW turbine ofinterest to the lab, the Polaris P21-60. But information on blade geometry (airfoils, chord length, twist) is
proprietary with no information available online [9]. Instead, blade design is based on the AOC 15/50 windturbine, a 50-kW turbine well-documented by the National Renewable Energy Laboratory (NREL). Information
on blade geometry is provided in Table 1 [2]. Twist is optimized for performance at a blade pitch of 4° by QBlade[4]. The resulting blade geometry is shown in Figure 4.
Table 1. Table of geometric airfoils over blade length (out from a hub of radius 0.305 m) used in QBlade simulation.
Position (%)
Chord (mm)Twist (°)
Airfoil
5469
60.7733S814
15
594
44.8478
S814
25
719
24.766S814
35
798
16.1283S814
45
738
11.5355S814
55677
9.71968S812
65
617
7.82523S812
75
557
6.46623S812
85
497
5.44484S813
95
437
5.51314
S813

Figure 4. AOC 15-50 turbine blade geometry in QBlade.

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Simulations of the blade were conducted under the range of conditions specified in Table 2. A plot of the resulting
Cp (λ, 𝛽) data is shown in Figure 6. This plot was compared to a sample Cp (λ, 𝛽) surface from NREL for adifferent wind turbine (the CART3 650-kW wind turbine) (Figure 5) for similarity in shape.
Table 2. Ranges of wind speed u, rotational speed ⍵r, and blade pitch 𝛽 for QBlade simulations.
Parameter
Range
Units
u
[0, 13]m/s[6, 20]rpm⍵r[-10, 10]°𝛽Figure 5. Cp-TSR-𝛽 surface for the CART3 60-kW windturbine [3].

Figure 6. Cp-TSR-𝛽 surface for the AOC 15-50 50-kW
wind turbine, generated in QBlade.

An existing Cp-TSR curve for an NREL 50-kW turbine at 0° is shown in Figure 7. When plotted on top of the
QBlade data shown in Figure 8, the Cp-TSR curve for the 50-kW NREL turbine sits close to the surface andfollows the same curvature, overshooting slightly at high TSR values.

Figure 7. Cp-TSR curves for the NREL 50-kW turbine
and the Polaris 60-kW turbine [5].

Figure 8. Cp-TSR curve for the NREL 50-kW turbine(blue line) and QBlade data for the AOC 15-50.

BLADE PITCH IMPLEMENTATIONGENERAL EQUATION FOR WIND TURBINE POWER COEFFICIENT
A general equation for modeling the power coefficient of wind turbine rotors is given in Eq. (6) [1]:𝑐𝐶𝑝 (𝜆, 𝛽) = 𝑐1 ( 2 − 𝑐3 𝛽 − 𝑐4 𝛽 𝑥 − 𝑐5 )𝑒 −𝑐6/𝜆𝑖
𝜆𝑖

1
𝜆𝑖

=

1
𝜆+0.08𝛽−0.035
𝛽 3 +1(6)(7)

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with constants c1 through c6 found empirically. x is assumed to be 2 for all calculations. Derived constants are
summarized in Table 3 [5].
Table 3. Constants used with Eq. (6) to model turbine performance, with source described.
Source
c1
c2
c3
c4
c5
Existing theoretical 60-kW
0.1564
116
0.4
0
3.5
Least squares 50-kW
4.86253.6654-0.85317.86090.1251
Scaled least squares 60-kW
4.15343.6654-0.85317.86090.1251c6
11.5
10.3240
10.3240

The constants used to produce the 60-kW curve found in Figure 7 were chosen by rescaling the 50-kW curve,
based on real data, for a 60-kW turbine, using the nominal maximum Cp of the P21-60 turbine [5]. Assuming avalue of c4 = 0 means that Cp does not vary significantly with blade pitch and is thus inaccurate for blade pitches
other than 0°. New constants that more accurately represented the relation between Cp and 𝛽 were needed.
Using the Cp-TSR-𝛽 plot generated for the AOC 15/50 in QBlade as a reference, the same methodology wasfollowed to determine new constants: fitting a surface in the form of Eq. (6) for a 50-kW turbine as closely as
possible, then scaling that equation for a 60-kW turbine. To achieve a closer fit between the approximate modeland the QBlade data, nonlinear regression was performed using MATLAB to minimize error between the modeland the data.
The first attempt at fitting the function to the data produced unrealistic results. The data generated by QBlade
included negative data points for Cp for large values of λ. After removing points from the QBlade data withnegative Cp values, least squares regression was repeated, yielding the constants found in Row 2 of Table 3. The
resulting 50-kW surface, compared with QBlade results and experimental NREL 50-kW results, can be seen in
Figure 9.

Figure 9. 50-kW Cp-TSR-𝛽 surface generated using Eq.
(6) (yellow surface) using constants found with leastsquares regression. QBlade AOC 15/50 results (dots) andthe NREL 50-kW Cp-TSR curve (blue) for 0° pitch
included for comparison.

Figure 10. 60-kW Cp-TSR-𝛽 surface generated using Eq.
(6) (yellow surface) based on scaled-down constants foundfor a 50-kW turbine. The old Cp-TSR curve (light blue) for
0° pitch is included for comparison.

To scale the 50-kW to match the NREL 60-kW reference curve, constant c1 for the 50-kW surface was divided
by the surface’s maximum Cp value and multiplied by the maximum Cp value of the NREL reference curve.
The scaled constants can be found in Row 3 of Table 3. The resulting 60-kW surface, compared with the old
Cp-TSR curve (light blue) for 0° pitch, can be seen in Figure 10. Because the QBlade design optimized for 𝛽= 4° and the old Cp-TSR curve was generated assuming optimum performance at 0°, comparisons to the oldcurve should be drawn at 4° rather than 0°.

SIMULINK IMPLEMENTATION
After deriving the necessary constants to characterize the 60-kW turbine, Eq. (6) was incorporated into the existing
Simulink model. To convert the calculated Cp to rotor torque, the following two additional relations were applied:1
𝑃 = 𝐶𝑝 ( 𝜌𝑣 3 𝐴)
(8)
2

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𝜏𝑟𝑜𝑡 =

𝑃𝜔𝑟(9)

where P is power generated, and A is the swept area of the rotor. To test the new system’s performance, step wind
inputs were fed to the old lookup table system (at 𝛽 = 0°) and the new system based on Eq. (6) (at both 𝛽 = 4° and
𝛽 = 0°). Results are plotted in Figure 11. System performance at optimum blade pitch is indistinguishable between
the old and new models.

Figure 11. Comparison of system outputs (rotor torque 𝜏rot [Nm], rotor speed ⍵r [rad/s] and line pressure phst [Pa]) for
stepped wind input v [m/s]. Results using Eq. (6) and the constants found using least squares regression at a blade pitch of 0°
and 4° are shown in yellow and red. Results using the lookup table torque calculation at a blade pitch of 0° are shown in blue
and are indistinguishable from those in red (using Eq. (6) and 𝛽 = 4°).

DYNAMIC PITCH CONTROL
Following implementation in Simulink, a controller for blade pitch was designed to limit power generation in
Region 3, when available power exceeds rated power for the turbine. Actuators for blade pitching were modeled
as a first-order system. Experimental time constants ranged between 9.6 and 12.3 seconds, so a time constant of
10 seconds was assumed. The blades were modeled as follows:1𝛽̇ + 𝛽 = 𝑓(𝑡)(10)𝜏𝛽where 𝜏𝛽 is the time constant for blade pitching and f(t) is the input to the system. To control blade pitch based on
power generation, a proportional controller comparing current power generation to rated power generation was
designed. The proportional controller’s output runs through a saturation block allowing only negative signals to
pass. If the turbine runs above rated power, blade pitch will drop, limiting power generation. But if the turbine isbelow rated power, the controller output will be optimal, that is, unchanged by the proportional controller. The
controller was tested with a 0.1 Hz sinusoidal wind input, yielding the response found in Figure 12. When overlaid
with the response of the uncontrolled pitch system locked at 𝛽 = 4°, the effect of the controller is evident; poweris limited to 60 kW by pitching the blades out of the wind when power would otherwise go above rated.

Figure 12. Overlaid comparison of Simulink outputs for system with (blue) and without (orange) pitch control.

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CONCLUSION
To improve the dynamic model of a 60-kW hydrostatic wind turbine, blade pitch was added as a variable in rotor
torque calculations. This capability was added to a previous MATLAB Simulink model of a hydrostatic windturbine with K⍵2 control. To incorporate blade pitching, a 50-kW wind turbine blade was recreated according todesign specs for the AOC 15/50 wind turbine and simulated using QBlade blade element momentum analysis.
Least squares regression was used to fit a general equation for power coefficient Cp to QBlade simulation results;
the resulting surface was scaled and compared to trusted experimental data. A proportional controller based on
pitch was added to the system to improve power capture and dissipate excess energy, modeling pitch actuators as
a first-order system. Further studies in improving wind turbine performance might explore the potential of using
stored energy using a hydraulic accumulator to pitch the blades and optimize power capture. There is also thepotential to use the simulation to study more sophisticated control methods.

ACKNOWLEDGMENTS
The authors thank the following members of the Hydraulic Power Transmission Laboratory at the University of
Minnesota for their help with this research: Justin Chen, Daniel Escobar, Mike Gust, Mark Leinberger,
Biswaranjan Mohanty and Eric Mohr.

REFERENCES
1. ACKERMAN, T., Wind Power in Power Systems, Second Edition, John Wiley & Sons, 2012.
2. HUGHES, P., SHERWIN, R., Advanced Wind Turbine Design Studies, Advanced Conceptual Study Final Report,
National Renewable Energy Laboratory, 1994.
3. JOHNSON, K. E., PAO, L. Y., BALAS, M. J., FINGERSH, L. J., Control of Variable-speed Wind Turbines: Standard andAdaptive Techniques for Maximizing Energy Capture. IEEE Control Systems Magazine, vol. 26, no. 3, 2006, pp. 70-81.
4. MARTEN, D., LENNIE, M., PECHLIVANOGLOU, G., NAYER, C. N., PASCHEREIT, C. O., Validation andApplication of an Unsteady Lifting Line Free Vortex Wake Module Implemented Within the Open-Source Wind Turbine
Blade Design Code QBlade, Journal of Engineering for Gas Turbines and Power, 138(7), 2015.
5. MOHANTY, B., “Design and Control of Hydrostatic Continuous Variable Transmission for a Community Wind Turbine,”
Ph.D. dissertation, University of Minnesota, Minneapolis, MN, USA, 2019.
6. MOHR, E., MOHANTY, B., STELSON, K. A., Short-Term Energy Storage System for Hydraulic Hybrid Wind Turbine
Transmission. ASME/Bath Symposium on Fluid Power and Motion Control, Bath, UK, 2020.
7. LAKS, J.H., PAO, L. Y., WRIGHT, A.D., Control of Wind Turbines: Past, Present, and Future, American Control
Conference 2009, St. Louis, USA, 2009.
8. PAO, L. Y., JOHNSON, K. E., A Tutorial on the Dynamics and Control of Wind Turbines and Wind Farms, American
Control Conference 2009, St. Louis, USA, 2009.
9. Polaris America LLC, P21-60 Specification Sheet, 2015.
10. STELSON, K. A., Hydrostatic Transmissions for Community-Scale Wind Turbines, 2021.

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Internal Flow and Hysteresis Characteristic of the Poppet Type
Pressure Control Valve
Seiei MASUDA*, Fumio SHIMIZU**, Masaki FUCHIWAKI**, Kazuhiro TANAKA***

Control Systems Engineering Department, IHI Corporation
3975-18, Haijima-cho, Akishima-shi Tokyo 196-8686 Japan(E-mail: masuda.seiei624@mail.kyutech.jp/masuda4361@ihi-g.com)**
Graduate School of Computer Science and Systems Engineering, Kyushu Institute of Technology
680-4, Kawazu, Iizuka-city, Fukuoka, 820-8502 Japan
Abstract. Hysteresis characteristics have emerged in pressure-flow characteristic experiments of poppet valves
designed and prototyped by the author. In order to clarify the mechanism of this hysteresis characteristic, we
confirmed the reproduction of hysteresis and confirmed the validity of the research policy. Specifically, weinvestigated the internal flow by CFD analysis and examined the reproduction of hysteresis characteristics by CFD.
First, the precision of CFD analysis was verified by confirming that the empirical values of in the reference andcalculated values were in good agreement. Furthermore, CFD analysis of the internal flow was performed. As aresult, it was clarified that there is a flow pattern that can be roughly divided into two types of internal flow withdifferent flow rates. Then, by changing the flow rate in a time-dependent manner using CFD, we were able toreproduce the hysteresis characteristics that are in good agreement with the empirical results.
Keywords: Poppet valve, Flow force, Hysterics, Pressure control valve, CFD

NOMENCLATUR
Ao
Ap
C
C0
Cd
Cf
D

k
： Opening area of damping orifice： Spring constantm： Cross-sectional area of poppet piston： Mass of the valve
Pressure ： Non-dimensional quantity of static pressure at： Orifice flow coefficientsurface of poppet valve： Critical damping coefficient(based on pressure at rated conditions)： Damping coefficient of the orifice
ΔP： Viscous damping coefficient： Differential pressure at the orificeQ： Damping coefficient in the equation of motion： Flow rateof the valve
Qo： Flow rate through the orifice
DB ： Hole diameter at seat
Radius ： Non-dimensional quantity of r
DL ： Valve piston diameter(based on inlet radius)
DS ： Diameter of poppet valve spool
Re： Reynolds numberdr： Spring displacement： Radial position
Fft： Flow force： TimeFlow rate ： Non-dimensional quantity of flow ratex
： Valve displacement(based on rated conditions)(based on valve displacement at rated
Fpreload ： Spring preload
conditions)
Fpressure ： Force applied to the valve due to pressurex0： Initial spring deflectionFsα： Spring force： Poppet cone angle
Fshear ： Shear force on the valveρ： Fluid densityg
ND ： Non-dimensional unit
： Non-dimensional quantity of x
(based on bore size)

INTRODUCTION
Poppet valves are the valves that are often used next to spool valves and are often used as pressure control
valves or flow control valves for hydraulic systems. Many studies have been conducted. Oshima et al. reported
the experimental results of the effect of valve shape on the characteristics of poppet valves [1]. They reported
the experimental results of a method to reduce the flow force of a poppet valve by changing the shape of thevalve and the shape of the roll downstream of the valve (hereinafter referred to as the collar). Ito et al. measured
the pressure distribution and flow force on the valve surface around a poppet valve and compared the resultswith numerical analysis [2]. Johnston et al. reported the experimental results of flow patterns and flow force in1
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poppet valves without and with collars. In the visualization experiment using a poppet valve without a collar in
the report, it was shown that there were three types of flow patterns and that they affected the flow force [3].
Vaugan et al. performed computational fluid dynamics (CFD) around a poppet valve using the finite volume
method with a commercial code to analyze the flow field [4], and Rundo et al. performed CFD on a poppet
valve used in a relief valve that can set the control pressure and predicted the flow pattern that generates theflow force. They reported on a method to reduce the flow force by using a collar [5]- [6]. Many studies havealso been conducted on the stability of poppet-type pressure control valves. Hayashi et al. reported that there are
two types of vibration modes depending on whether the poppet is in contact with the valve seat or not [7].
However, no clear stability criterion for poppet-type pressure control valves has been obtained. The motivation
for this study was that the pressure-flow characteristics of a prototype poppet valve [10] designed by the author
showed a hysteresis characteristic in which the polarity of the pressure gradient changed with respect to the flowrate change with a very long period [10]. This hysteresis characteristic has not been studied before, and toclarify the mechanism of its occurrence, we studied the internal flow by CFD analysis and the reproduction of
its characteristics by CFD. The hysteresis phenomenon is one of the typical nonlinear phenomena, and it cannot
be denied that it may have a significant influence on the stability of poppet valves and poppet-type hydrauliccontrol valves.
Figure 1 shows the geometry of the prototype pressure control valve, and Table 1 shows its non-dimensionalized
main parameters. Figure 2 shows the hysteresis characteristics of the pressure-flow characteristics obtained in
the experiment. The arrows (1) to (5) in Figure 2 show the time sequence of the control pressure trajectory.
When the flow rate reaches about 80% of the rated value, the pressure in path (3) appears to increase rapidly at aconstant flow rate. When the mass flow rate reaches about 80% of the rated condition, however, the pressure inpath (3) seems to increase rapidly at a constant flow rate. Even if the flow rate is decreased in the return path
after reaching the maximum flow rate, the pressure remains high and decreases gradually. Therefore, thepurpose of this study was to reproduce the hysteresis phenomenon of the pressure-flow characteristics byanalyzing the internal flow of the test poppet valve using CFD.

Table 1. Specifications of Pressure Control valve

Constants, Dimensions and
Rated Performance
Poppet apex angle(α) [degree]
Large spool diameter ratio(DL/DB)
Spool small diameter ratio(DS/DB)
Valve seat diameter (DB/DB)
Resonant frequency [rad/s]

Value
90
1.8
1.07
1
800

Note 1: Re = (Mean velocity at throttle･Distance between
poppet and seat) / Kinematic viscosity of working fluid

Figure 1. Structure of Pressure Control Valve

(3)

(4)
(2)

(5)

(1)

Figure 2. Control Pressure vs. Flow Rate Hysteresis Characteristic2

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VALIDATION OF ANALYSIS METHOD
At the beginning of the study, we verified whether the CFD analysis of the flow field around the poppet valve
could reproduce the real phenomena. The test data used for the verification is the data of Ito et al. [2]. The
commercial code Simerics-MP+ is used to analyze the flow field of an incompressible fluid using the finite
volume method. Simerics-MP+ is capable of CFD analysis and coupled analysis of the equations of motion of
the same body. By applying a sliding mesh to the finite volume method [8], the CFD analysis can reflect thetransient changes in valve opening and flow path geometry as the valve is displaced. Figure 3 shows thegeometry of the poppet valve of Ito et al. used in the CFD analysis. Figure 4 shows the test apparatus. Figure 5
shows the analysis grid. The mesh type is an orthogonal grid with 8 to 10 grids at the throat, and the minimum
mesh size is on the order of 1/100 mm. No special consideration is given to the boundary layer mesh and wallroughness.
Poppet Surface Pressure
Measurement Point

Inlet

OutletPoppet Valve Shape

R(=18)

Measurement point:

r1
Radius=r1/R
Measurement Point

Definition of Non-Dimensional RadiusFigure 3. Shape of Poppet Valve used for Verification of Analysis Method

Table 2. Analysis Conditions for a Simerics-MP+ Model

Software used
Analysis type

Simerics-MP+
Steady state

Turbulence model

standard k-ε
3

Fluid density [kg/m ]Fluid viscosity [Pa・s]
Entrance boundary condition[l/min]Exit boundary condition [Pa]Valve displacement [mm]Grid type863
0.0267(a) Q = 42 (Re = 200)(b) Q = 60 (Re = 300)(c) Q = 42 (Re = 200)0 (Static Pressure)(a) 6, (g = 1/6) (1)(b) 6, (g = 1/6)(c) 4, (g = 1/9)
Orthogonal grid

Grid number
1,300,000 [-]Note1: g = Valve displacement / Inlet pipe diameter

Figure 4. Verification Test Setup

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Table 2 shows the three conditions used for verification. Although a cavitation model was incorporated in the
CFD calculations, the maximum calculated void fraction was less than10-5, so the effect of cavitation was not
significant. We have also confirmed that the calculated results do not differ depending on the location of theoutlet boundary and the conditions. A comparison of the static pressure distribution on the surface of the poppet
valve between the test results and the CFD analysis is shown in Figure 6. In this figure, the pressure is nondimensionalized by the valve tip pressure and the radial direction is non-dimensionalized by the inlet port radius.
The calculated results are in good agreement with the experimental results of Ito et al. under three different
analytical conditions, and the difference between them is small, less than a few percent. In the next section, we
attempt to apply the Simerics-MP+ analysis to the flow field around the poppet valve with hysteresis as shownin Figure 1.

Whole Grid

Cross-Sectional ShapeFigure 5. Analysis Grid for Simerics-MP+(a) Q = 42(c) Q = 42(b) Q = 60
Figure 6. Verification Results

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ANALYSIS OF THE FLOW FIELD AROUND A POPPET VALVE ANDDISCUSSION
To investigate the flow pattern of the internal flow, a steady-state flow analysis was conducted with the flowrate at the inlet and the pressure (back pressure) at the outlet. As an example of the results of the steady-stateflow analysis, the calculation results at the four points shown in Table 3 are illustrated below. Figure 7 shows
the analysis grid. The CFD results (velocity contour of the poppet valve cross-section) for conditions (1) to (4)in Table 3 are shown in Figure 8.

Inlet

OutletAnalysis Model

Poppet Valve

Cross Section of Grid

Detailed around the Poppet

Figure 7. Analysis Grids for Simerics-MP+
Table 3. Analysis Conditions for a Simerics-MP+ Model
Software usedSimerics-MP+
Analysis type
Steady / Transient
Turbulence modelStandard k-ε

3

Fluid density [kg/m ]Fluid viscosity [Pa・s]
Entrance boundary condition [ND]
Exit boundary condition [Pa]7700.00149
Steady:(1) 0.30 (2) 0.49 (3) 0.73 (4) 1.00
Transient: 0.05 - 1
0.79 (Static Pressure)

Grid number

4,000,000 [-]Incremental time

- / 0.000054 [s]Condition (1) Flow rate [ND]=0.30,(g = 0.024, Re = 13000)
Condition (2) Flow rate [ND]=0.49,(g = 0.039, Re = 22000)
Condition (3) Flow rate [ND]=0.73,(g = 0.058, Re = 34000)
Condition (4) Flow rate [ND]=1.00,(g = 0.080, Re = 49000)
Figure 8. Velocity Contours and Flow Patterns under the Conditions shown in Table 7

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Figure 9. Distributions around the Poppet Valve

When the flow rate at the inlet boundary condition in Table 3 is small (Condition (1)), the velocity vector
diagram shows that the jet around the poppet valve flows along the poppet axis and the jet is attached to thespool (hereafter referred to as attaches flow). The jet then decelerates, collides with the spool collar, and turns at
a right angle. In Condition (2), (3) and (4), as the flow rate increases, the jet separates near the bottom of theminimum throat of the valve and flows in a straight line toward the port of the sleeve (straight flow). These
results indicate that there are two types of flow patterns in the valve: Attached flow and Straight flow.
Figure 9 shows the static pressure distribution at the valve surface for an arbitrary cut plane relative to thepoppet valve radial direction. The inlet pressure increases with increasing flow rate due to the increase inpressure at the apex of the poppet valve. In the radial pressure distribution, the pressure gradient changes near
the minimum throttle, and there is no significant difference in the large diameter of the valve. Comparing thepressure at a certain point of radius 0.7 ~ 1.0 with the increase of inlet flow rate for each flow rate condition, thepressure tends to be lower for the condition with higher inlet flow rate. The force in the direction of opening thevalve decreases. This decrease in force is equal to the increase in force in the direction of closing the valve.
Further elaboration is needed in this area.
Figure 10 shows the flow force against the valve displacement. In conditions (2) to (4), the flow force is
proportional to the valve displacement, but in condition (1), the flow force is calculated to be small, suggestingthat there is a correlation between the flow force and the flow pattern.
When the flow force acts on the valve, the inlet pressure changes due to the change in valve displacement andspring force, and the force due to pressure acting on the valve also changes. The flow force also changes. In
other words, the effect of the variation of the force acting on the valve is calculated by the coupled analysis
function of Simerics-MP+, and the final balanced spring displacement, inlet pressure, and flow force arecalculated at each time. In this way, the force acting on the valve is fed back by the flow force variation, and thevalve displacement and inlet pressure are recursively determined. Since the opening area of the poppet valve is
proportional to the valve displacement as shown in Equation below [9], the flow force is almost proportional tothe valve displacement except for condition (1). The polarity acts in the direction of closing the valve. In
condition (1), the flow force is much smaller than the other conditions, and the flow force per displacement isless than 1/5.

Ap =   DB  x  sin


2

(1)

Figure 11 shows the valve opening area versus valve displacement, and the results show that the valve opening
area is proportional to the valve displacement. Figure 12 shows the calculated flow coefficients for eachcondition, where the opening area is non-dimensionalized as the opening area under the rated condition. As canbe seen from these results, the flow coefficient is constant at about 0.75 regardless of the change in flow rate,
indicating that the state of attachment and detachment of the jet does not affect the flow coefficient.
Next, the results of the steady-state flow analysis for the valve with hysteresis characteristics under the boundary
conditions (1) to (4) shown in Table 3 are shown in Figure 13, in comparison with the test results.

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Figure 10. Valve Displacement vs. Flow Force

Figure 11. Valve Displacement vs. Valve Opening Area

Figure 12. Coefficient of the Flow Rate

Figure 13. Control Pressure Flow Rate Hysteresis
Characteristics of the Pressure Control Valve(Empirical results)

From this comparison, the analytical results agree well with the characteristics of the high-pressure side at thereturn of the flow characteristics against the control pressure with hysteresis. This point is interesting, but thedetails of the reason for this are still unknown at this stage, and further research is needed.

RESULTS AND DISCUSSION OF COUPLED ANALYSIS OF VALVEEQUATIONS OF MOTION AND CFD
To reproduce the hysteresis of the flow characteristics against the control pressure of a poppet valve, a coupledanalysis between the flow field of the pressure and flow characteristics and the equation of motion of the valve
was conducted using Simerics-MP+. The equation of motion of the valve is expressed as Eq. (2) [9].

m

d 2x
dx
+D
+ k d =  FPressure +  FShear − FPreload
2
dt
dt

(2)

The external forces acting on the valve are the force due to the pressure to which the valve is subjected, theshear force and the spring preload. In each time increment of the CFD calculation, when the convergencejudgment of the flow calculation is completed, the calculation module of the above equation of motion of thevalve is activated and the translational motion of the valve is calculated. The damping coefficient D in the above
equation of motion can be the sum of the damping coefficient Cf due to the viscosity of the fluid and thedamping coefficient Cd due to the damping orifice located behind the valve and can be expressed as “D = Cf +Cd”. First, we consider the damping force of the damping orifice. The damping force of the orifice as a valve
circuit can be equivalent to the force corresponding to the pressure drop ΔP at the orifice as Eq. (3).

Cd 

dx
= P  AP
dt

(3)

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From this, it is possible to calculate the damping force. Since ΔP changes transiently in experiments, it isdifficult to determine it exactly, but the damping force can be calculated by referring to the value in theexperiment. Furthermore, the damping coefficient of the damping orifice is determined. Since the valve velocity
is the flow rate through the orifice divided by the poppet piston area, the damping coefficient is can becalculated from Eq. (3) and Eq. (4). The valve velocity can be estimated.
1

dx Qo Ao C  2  P  2
=
=


dt AP
AP   (4)

Next, we consider the damping force due to the viscosity of the fluid. The largest of these is the viscous
frictional force between the outer circumference of the piston, which is the large diameter part of the valve, andthe sleeve surface, which is derived from Newton's law of viscosity. This value is less than 1/10 of the damping
force generated by the orifice, according to the results of calculations. These results show that the viscousdamping force is small compared to the damping force of the damping orifice, indicating that this orifice has asignificant influence on the actual damping of the poppet valve. The calculation method of the axial flow force
is shown in Figure 14.
Figure 15 shows the temporal variation of the supply flow rate (ramp input). In the experiment, the data wascollected at a time that could be considered steady state, but in the CFD calculation, the same real time
calculation as in the experiment is not possible due to the condition of the Courant number. Therefore, a largerchange in inlet flow rate than in the experiment was input to the calculation.
A comparison between the simulation results and the test results of the flow rate against the control pressure of
the pressure control valve is shown in Figure 16. Because of the large computational load of this analysis, thesimulation results at the end of the forward journey (under rated conditions) were used as initial values for the
return journey. This figure shows the calculation results after the transient condition that occurs at the start ofvalve opening is settled. The control pressure-flow characteristics, including the hysteresis phenomenon, are ingood agreement. Compared with the pressure characteristic curve with empirical hysteresis shown in Figure 16,
the numerical results of the forward path with slowly increasing flow rate were almost equal to the empiricalpressure value. The numerically calculated pressure was almost equal to the value of the curve for the lowpressure side. The velocity contour diagram was checked for the flow conditions from (1) to (4) in Table 3, andit was found that attached flow occurred in conditions (2) and (3) as in condition (1), and flow force wasworking in the direction of opening the valve. On the other hand, for the return path, the results of the numericalanalysis and the empirical pressure characteristic curve for the high-pressure side are in good agreement in order.
And the results are comparable to those calculated under the Steady state condition shown in Figure 13. The
flow pattern is strait flow with the jet detached after passing through the minimum opening area of the poppet.
The estimated value of steady-state flow force and the flow force obtained by CFD are in good agreement inorder. The flow force in forward path is on order the same as the flow force in return path, but the polarity isreversed.

Axial force acting on the valve body
In case of valve closure
In case of valve release
Axial flow force
Figure 14. Calculation way of flow force

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Figure 15. Input Conditions for supply flow rate

Figure 16. pressure flow Characteristics

Figure 17. Comparison of Flow Force Characteristics

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Figure 17 shows a comparison of the flow forces at six points on the forward path calculated from the results ofthe coupled analysis, the flow forces at four points calculated under the conditions (1) to (4) in Table 3, and thefour calculated values of the steady-state flow force based on the momentum theory at that time. The estimatedvalues of the steady-state flow force and the CFD-derived flow force agree well in terms of order. The order ofthe flow force in the forward path is the same as that of the flow force in the return path, but the polarity isreversed. The fact that the hysteresis characteristic can be reproduced in the calculation despite the large
difference in the flow input condition dQ/dt between the experiment and the numerical calculation may meanthat this hysteresis characteristic always occurs regardless of the length of the period. Alternatively, it may meanthat the time required for this phenomenon to reach a quasi-steady state is extremely short. Since the damping
coefficient is also relevant here, further research is needed.

CONCLUSION
Hysteresis characteristics appeared in the characteristic tests of the prototype poppet valve. It was found that thehysteresis characteristic appeared depending on the operating conditions and valve shape. CFD analysis wasused to elucidate the mechanism of this characteristic.
First, CFD analysis of the flow field around the poppet valve was performed using a commercial code, and thevalidity of the analysis results was confirmed by comparing them with test results. The analysis of a poppet-type
pressure control valve, which generates hysteresis in the control flow characteristics against the control pressure,revealed that the flow field around the poppet valve can be classified into two types: attached flow and straightflow. It was found that the flow field around the poppet valve can be divided into two types: adhered flow anddetached flow. It was also found that the hysteresis phenomenon that occurs in the control flow characteristic of
the poppet-type pressure control valve against the control pressure can be reproduced by the coupled analysis of
the CFD calculation of the total pressure around the poppet valve and the equation of motion.
This paper is an English version of the paper [10] submitted by the authors to the Journal of the JFPS 2021 forthe 11th JFPS International Symposium.

ACKNOWLEDGMENTS
We would like to express our deep gratitude to Mr. Naoki Hirose (currently working at Voith IHI Paper
Technology Co., Ltd.), who was a master's student at Kyushu Institute of Technology, for his cooperation in thisresearch.

REFERENCES
1.
2.
3.
4.
5.
6.
7.
8.
9.
10.

OSHIMA, S., AN EXPERIMENTAL STUDY ON THE SEVERAL POPPET VALVES WITH DIFFERENCE IN
SHAPE, Proceedings of the JFPS International Symposium on Fluid Power, 1989, pp.361-368
ITO, K., TAKAHASHI, K., INOUE, K., Flow in a Poppet Valve: Computation of Pressure Distribution using a
Streamline Coordinate System, JSME International Journal Series B Fluids and Thermal Engineering, 1993, 36-1,
pp.42-50
JOHNSTON, D. N., EDGE, K. A., Experimental Investigation of Flow and Force Characteristics of Hydraulic Poppet
and Disc Valves, Proceedings of the Institution of Mechanical Engineers, Journal of Power and Energy, 1991, 205-3,
pp.161-171
VAUGHAN, N. D., JOHNSTON, D. N., EDGE, K. A., Numerical simulation of fluid flow in poppet valves,
Proceedings of the Institution of Mechanical Engineers, Journal of Power and Energy, 1992, 206-2, pp.119-P127
ALTARE, G., RUNDO, M., OLIVETTI, M., 3D Dynamic Simulation of a Flow Force Compensated Pressure Relief
Valve, ASME International Mechanical Engineering Congress and Exposition, 2016, pp. V007T09A061
FINESSO, R., RUNDO, M., Numerical and experimental investigation on a conical poppet relief valve with flow force
compensation, International Journal of Fluid Power, 2017, 18-2, pp.111-122
HAYASHI, S., Instability of Poppet Valve Circuit, JSME international journal. Ser. C, Dynamics, control, robotics,
design and manufacturing, 1995, 38.3: 357-366.
Simerics Manual V5.0, 2019, Simerics, Washington
New Edition hydraulic and Pneumatic Handbook, Ohm Sha, Tokyo, 1989, pp.69-71 (Book in Japanese)
MASUDA, S., SHIMIZU, F., FUCHIWAKI, M., TANAKA, K., Internal Flow and Hysteresis Characteristic of the
Poppet Type Pressure Control Valve, Journal of the JFPS, 2021, 52-1, pp.8-15.

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Flow Patterns and Hysteresis Characteristic of a Poppet Valve
Naoki HIROSE*, Seiei MASUDA **, Fumio Shimizu***,Masaki Fuchiwaki***, Kazuhiro TANAKA****

Paper Machinery Engineering Department, Paper Machine Group,
Voith IHI Paper Technology Co., Ltd.
2-1-6 Tsukuda, Chuo-ku, Tokyo 104-0051, Japan(E-mail: naoki.hirose@voith.com)**
Control Systems Engineering Department, IHI Corporation
3975-18, Haijima town, Akishima-shi, Tokyo, 196-8686 Japan(E-mail: masuda4361@ihi-g.com)*** Graduate School of Computer Science and Systems Engineering,
Kyushu Institute of Technology,
680-4, Kawazu, Iizuka city, Fukuoka, 820-8502 Japan

Abstract. Previous studies on poppet valves have revealed that there are various flow patterns in the flow fieldaround the valve body and that the flow pattern inside the poppet valve has various effects on the poppet valve.
Recently, Masuda et al. have reproduced the hysteresis phenomenon on the pressure and flowrate characteristics of apoppet valve by CFD analysis coupled with an equation of motion. Unfortunately, their paper does not describe thedetailed relationship between the hysteresis characteristics and the flow pattern. In this study, by succeeding anddeveloping the work of Masuda et al. and including the results of previous studies, we have systematically organized
the flow patterns of poppet valves into five types and clarified the relationship between the hysteresis characteristics
and flow patterns. In addition, the flow force acting on the poppet valve connecting the hysteresis characteristic andthe flow pattern has also been clarified.
Keywords: Poppet valve, Flow patterns, Flow forces, Hysteresis characteristics, CFD

INTRODUCTION
Poppet valves are the second most commonly used control valves in hydraulic systems after spool valves, andhave been widely studied [1]. It is well known that the flow pattern in a poppet valve shows more complex
changes than that in a spool valve and can be one of the factors that affect the dynamic characteristics of thevalve. Masuda et al. [2] showed that the hysteresis phenomenon of the pressure-flow rate characteristics of apoppet valve can be reproduced by calculating the integral of the valve driving force due to the pressure aroundthe poppet using CFD analysis coupled with the equation of motion. Figure 1 shows the geometry of the
prototype pressure control valve, Figure 2 shows the experimental circuit, and Figure 3 shows the experimental
and CFD results of the pressure-flow rate characteristics. The results of this study are interesting because theyshow that CFD software can reproduce the hysteresis characteristics of poppet valves. Unfortunately, their study
does not describe the detailed relationship between the hysteresis and the flow pattern.
In many previous studies on flow patterns, Schrenk [3] first visualized the flow pattern under the same valve
opening, the same supply pressure, and the same flow rate, and reported the existence of five types of flow
patterns. Figure 4 shows the schematic diagram of the five flow patterns he showed. Following this, McCloy [4]observed flow patterns in a two-dimensional channel and identified three types of flow patterns. Johnston et al.
[5] reported the results of experiments on flow patterns and flow forces in poppet valves with and without
collars. In the visualization experiment using a poppet valve without a color, they reported the existence of three
types of flow patterns, two of which (Attached flow pattern and Straight flow pattern) occurred stably. They also
showed that these flow patterns affect the flow force, and reported that the Attached flow pattern generates anegative flow force and the Straight flow pattern generates a positive flow force. Figure 5 is their visualization
resultant diagram showing those three flow patterns. Furthermore, Vaugan et al. [6] performed a CFD analysisof the internal flow of a poppet valve using axisymmetric incompressible flow, steady flow analysis, k-ε
turbulence model, and finite volume method using CFD software. They reported that the quantitative predictionof jet reattachment was inaccurate and the Attached flow pattern could not be reproduced. Hayashi [7]summarized the unstable vibration generated in poppet valves based on these studies. Rundo et al. [8] performed
CFD analysis of the flow inside a poppet valve used in a relief valve that can set the control pressure, predicted
the flow pattern that generates the flow force, and reported a method to reduce the flow force with a collar. Thus,
it is understood that there are various flow patterns in the flow field inside the poppet valve and that they are
related to the unstable vibration, but a systematic understanding of them has not yet been obtained.

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In order to succeed and further develop the work of Masuda et al. [2], this study systematically organizes theflow patterns in the prototype poppet valve used in their research by CFD analysis, and clarifies the relationshipbetween the hysteresis characteristics and the flow patterns, as well as the relationship between them and theflow force. Simerics-MP+ was used as the CFD software.

Figure 1. Structure of Pressure Control Valve [2]Figure 2. Schematic Diagram of Experimental Equipment [2]Figure 3. Hysteresis Path in Control Pressure vs. Flow Rate
(Comparison of Experimental and Transient CFD Results) [2]2

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Figure 4. Flow Patterns from Schrenk [4]Figure 5. Flow Patterns from Johnston [5]NOMENCLATUR
D
DB
DL
DS
F
FshearFpreload
Fpressure
k
m
ND
P
Q
t
x
α
ρ
μ

： Damping coefficient [N/(m/s)]
Sheet section hole diameter：(= 13.5 [mm])： Valve piston diameter [mm]： Poppet valve spool diameter [mm]： Flow force [ND]
： Shear force [N].
： Spring preload [N]： Pressure force [N]： Spring coefficient [N/m]： Poppet valve mass [kg].
： Dimensionless unit
： Pressure [ND]
： Flow rate [ND]： Time [s]： Valve displacement [ND]
： Poppet cone angle (= 45 [°])： Density [kg/m3]
Viscosity [Pa・s]

CALCULATION CONDITIONS FOR THE FLOW FIELD
INSIDE A POPPET VALVEIn Simerics-MP+, the equations of motion of the valve body can be coupled with CFD analysis. The
computational grid is a sliding mesh that can expand and contract to accommodate the translational motion of
the valve. By applying the finite volume method to the sliding mesh, CFD analysis can be performed to reflect
the transient changes in the valve opening degree and the flow path shape with the valve disc displacement.
Figure 6 shows the schematic of the test poppet valve, Figure 7 shows the calculation grid, and Table 1 shows
the calculation conditions.
In this study, steady flow and unsteady flow calculations were performed. In the steady flow calculation, thevalve was fixed and the condition of constant flow rate was applied. In the unsteady flow calculation, the valve
was made movable and the flow rate was increased linearly in the range of Q = 0.05~1.0 [ND], and when the
rated flow rate was reached, the flow rate was decreased linearly in the range of Q = 1.0~0.05 [ND].
For the initial conditions of the unsteady flow calculation, the initial value of the flow rate was first determined
to be Q = 0.05 [ND], and then the unsteady flow calculation was carried out with the valve body set to be3

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movable, and the valve displacement when the valve motion was settled down was used as the initial value. The
valve displacement at this time was x = 0.04 [ND]. The pressure P, flow rate Q, and valve displacement x used
are nondimensionalized in terms of rated pressure, rated flow rate, and valve displacement at rated pressure andflow rate, respectively.

Figure 6. Outline Bird's-Eye and Cross-Section Viewof the Test Poppet Valve

Figure 7. Analysis Grid of the Target

Table 1. Analysis Conditions

Software usedSimerics-MP+

Analysis type

Steady stateTransient

Turbulence model

standard k-ε

Fluid density

770

Fluid viscosity

0.00149

Entrance boundary
condition [ND]
Exit B.C. [ND]

0.05~1(Flow rate/Rated
flow rate )0.2 (Static Pressure)

Flow rate= const.

Grid number

4,000,000 [-]Incremental time

0.000054 [s]ANALYSIS RESULTS
Flow Patterns
Figure 8 shows five typical flow patterns obtained from steady-state calculations with a fixed valve and constantflow rate (for example, for a valve displacement x = 0.07 [ND]). Each figure shows the velocity and pressure
contour plots for the central cross section including the poppet axis. The flow patterns 2 to 5 were observed evenin the case of unsteady calculation under the condition that the valve was movable and the flow rate waschanging. In the present study, the initial valve displacement was set to 0.04 [ND] due to the conditions thatmust be satisfied in order to keep the calculation grid sound. The areas circled by the solid black lines in Figure
8 (a), (b), (d) are enlargements of the pressure distribution in the areas circled by the dashed black lines.
From these results, the flow patterns can be roughly classified as follows.
(a) Flow pattern 1
This is a viscous flow pattern that spreads uniformly through the channel at very low flow rates. Since it appears
only at very low flow rates, it does not have a significant effect on the flow in the valve in actual use, such as
flow force.
(b) Flow pattern 2
In the low flow rate region, the first point where the jet from the throat separates from the poppet surface is the4

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sleeve corner that constitutes the throat. Because the jet is attracted to the low pressure part of the separated flow
region, the jet adheres to the vertical wall of the sleeve inlet. This is a stable flow pattern.
(c) Flow pattern 3This is an unstable flow pattern that occurs in the low to medium flow rate region. Since the steady flow
analysis did not converge, the unsteady flow analysis was conducted. As a result, it was found that the jet flow
pattern oscillates between flow patterns 2 and 4, and its frequency is not constant.
(d) Flow pattern 4 (Attached flow)
This is a stable flow pattern that occurs in the medium to large flow rate region. The flow passes through thethroat and flows into the valve chamber as a jet, generating a flow that adheres to the surface of the poppet spool
due to the Coanda effect. If a collar is not present, the jet flows directly downstream to the outlet. If a collar ispresent, the jet hits the collar and turns 90 degrees in the radial direction.
(e) Flow pattern 5 (Straight flow)
This is an extremely stable flow pattern that occurs in the medium to large flow rate region. The flow passes
through the throat and flows into the valve chamber as a jet, which detaches from the end of the poppet cone andflows directly into the valve chamber as a free jet.
The characteristics of the above five flow patterns are illustrated in Figure 9. The numbers in the figure
correspond to the numbers of the flow patterns in Figure 8. Table 2 shows the results of the comparison between
the flow patterns shown in Figures 4 and 5 and those obtained from the present analysis.
As for the flow patterns in the poppet valve, there were some research results such as five types or three types offlow patterns from past visualization experiments, but each of them corresponds to one of the five flow patterns
shown here. Therefore, based on the numerical results of this study, the flow field in the poppet valve can besystematically classified into five types. Furthermore, flow patterns 4 and 5 are confirmed in all experimental
reports and computational results, and these two types of flow patterns can be recognized as stable flows. The
hysteresis characteristics in this study are strongly related to the regions of flow pattern 3, 4 and 5.

(a) Flow pattern 1(b) Flow pattern 2(c) Flow pattern 3 (Unstable flow)(d) Flow pattern 4 (Attached flow)(e) Flow pattern 5 (Straight flow)
Figure 8. Pressure and Velocity Contour Plots for Five Typical Flow Patterns in Poppet Valve(Obtained from Steady-State Calculations with a Fixed Valve and Constant Flow Rate)

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Table 2. Correspondence Table between Type
Number of Flow Patterns Calculated in
the Present Work and Flow Patterns from
Schrenk [3] and Johnston [5], where
Alphabets such (a)-(h) correspond to thatshown in Figures 4 and 5
Flow pattern
Present work

1

2

3

4

5

Schrenk

A

B

D, E

C

JohnstonA

C

B

Figure. 9 Illustration of the Calculated Flow Patterns

Switching Phenomena
In this section, the cause of the transient switching phenomenon from flow pattern 4 to 5 in the trajectory of thehysteresis characteristics shown in Figure 3 is examined on the reference planes ① and ② shown in Figure10(a), (b). The reference plane ① is set as a vertical cross section that includes the axis of the poppet valve, andthe reference plane ② is set as a circular cross section that cuts the separation bubble occurring from the end of
the poppet cone to the corner of the poppet spool start.
Contour plots of the eight flow fields (velocity distribution and two types of pressure distribution at the sametime), which characteristically suggest the switching mechanism of this hysteresis characteristics, are shown in
Figures 11, 12, and 13. Figure 11 is a contour plot of the velocity distribution of the internal flow on thereference plane ①, showing the jet flowing from the throat into the valve chamber. However, the color contoursof the poppet cone surface and the poppet spool surface in Figure 11 show the pressure distribution on thosesurfaces, respectively. Figure 12 shows the pressure distribution contours of the entire valve chamber on thereference plane ①. Figure 13 shows the pressure distribution contours on the reference plane ②. The following
alphabetic symbols in (a) ~ (h) correspond to the symbols in Figures 11~13. Summarised by time series, we
comment on the results as follows.
(a) t = 0.391 [s].
Both the flow pattern 4 and the pressure distribution are axisymmetric and stable.
(b) t = 0.427 [s].
Flow pattern 4 exists stably as an axisymmetric flow, but the pressure distribution at the reference plane ② isstarting to become unstable in terms of axisymmetry with the appearance of a low-pressure part on the left side.
Unfortunately, the reason for this remains unexplained.
(c) t = 0.433 [s].
The pressure distribution at the reference plane ② shows that the lower half of the cross section is under low
pressure due to the large growth of the low pressure section that had appeared on the left side, and theaxisymmetry of the pressure has broken down. The flow pattern of axisymmetric flow 4 is about to detach fromthe lower part of the valve spool.
(d) t = 0.435 [s].
The low pressure area that grew on the lower side of the pressure distribution in the reference plane ② above
repeats its strength and weakness, and is moving downward to the region closer to the outlet port. Due to thiseffect, the flow pattern 4 of the axisymmetric flow is trying to leave the lower part of the poppet spool. In otherwords, the transition to flow pattern 5 has begun.
(e) t = 0.448 [s].
The low pressure part of the pressure distribution in the reference plane ② repeatedly becomes larger andsmaller, mainly at the bottom. Due to this effect, flow pattern 4 exists only in the upper part of the poppet spool,
while the rest of the region is in a separated state (flow pattern 5).
(f) t = 0.461 [s].
The low pressure zone grows larger and expands to the upper half of the poppet spool. As a result, the flow
pattern on the upper half of the poppet spool also begins to separate. The flow pattern on the rest around the6

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poppet spool becomes completely separated and flow pattern 5 appears.
(g) t = 0.463 [s].
The low pressure part of the pressure distribution expands to the entire circumferential area, and the flow iscompletely separated, completing the flow pattern 5.
(h) t = 0.471 [s].
The flow pattern 5 is complete, and both the flow pattern and the pressure distribution are axisymmetric andstable.
In (a) ~ (h), there are two stable flow patterns 4 and 5, such as Attached flow and Straight flow. As the flow rateincreases, there is a transition (switching) between the two flow patterns. This is found to be due to the collapseof the axisymmetry of the flow as a result of the collapse of the axisymmetry of the pressure distribution of theseparated bubble (low pressure part) generated in the reference plane ② as the flow rate increases. These results
indicate that there is a close relationship between the path of the hysteresis characteristic and the flow pattern.

Reference direction

Reference direction(a) Plane ① (red surface)(b) Plane ② (green surface)vertical cross section to include the axis of the poppet valve
Circular cross section to cut separation bubbles
Figure 10. Reference Planes to Examine the Causes of Transient Switching from Flow Pattern 4 to 5(a) t = 0.391 [s](b) t = 0.427 [s](c) t = 0.433 [s](d) t = 0.435 [s](e) t = 0.448 [s](f) t = 0.461 [s](g) t = 0.463 [s](h) t = 0.471 [s]aFigure 11. Velocity Contours in the Plane ① (Pressure Contour on the Poppet Surface)
Time Series Data of the Phenomenon of Transient Switching from Flow Pattern 4 to 5

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The 11th JFPS International Symposium on Fluid Power 2020(a) t = 0.391 [s](b) t = 0.427 [s](c) t = 0.433 [s](d) t = 0.435 [s](e) t = 0.448 [s](f) t = 0.461 [s](g) t = 0.463 [s](h) t = 0.471 [s]Figure 12. Pressure Contours in the Plane ①
Time Series Data of the Phenomenon of Transient Switching from Flow Pattern 4 to 5(a) t = 0.391 [s](b) t = 0.427 [s](c) t = 0.433 [s](d) t = 0.435 [s](e) t = 0.448 [s](f) t = 0.461 [s](g) t = 0.463 [s](h) t = 0.471 [s]Figure 13. Pressure Contours in the Plane ②Time Series Data of the Phenomenon of Transient Switching from Flow Pattern 4 to 5

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FLOW FORCES ON AXIAL DIRECTION
Coupled Analysis between CFD and Equation of Motion of Valve Body
In order to reproduce the hysteresis characteristics of the flow rate against the control pressure in a poppet valve,a coupled analysis between the flow field of the pressure and flow rate characteristics and the equation of
motion of the valve body was conducted using Simerics-MP+. The equation of motion of the valve body in theaxial direction is expressed as follows: Eq. (1) [9].

m

d 2x
dx
D
 k d   FPr essure   FShear  FPr eload
2
dt
dt

(1)

The force acting on the valve body is expressed as the sum of the force due to the pressure, shear force andspring preload that the valve body receives.
In the coupled analysis, within each time increment of CFD calculation, after the convergence judgment of theflow calculation is completed, the module for calculating the above equation of motion of the valve body isactivated, and the translational motion of the valve body is calculated.
The damping coefficient D in the above equation of motion can be thought of as the sum of the damping
coefficient Df, which is due to the viscosity of the fluid, and the damping coefficient D d, which is due to thedamping orifice located behind the valve, and can be expressed as D = D f + Dd. Comparing the magnitudes of
the two, the damping force due to the viscosity of the fluid is an order of magnitude smaller than the damping
force of the damping orifice. That is, Df << Dd. These results indicate that this orifice has a significant influenceon the actual damping of the poppet valve. Since the critical damping coefficient is C0 = 2= 108 [N/(m/s)] , weconsider that a larger value would be acceptable as a setting value in the calculation. Therefore, we set thedamping coefficient to 2000 [N/(m/s)], referring to the value that is the smallest case of damping force.

Relationship between Flow Pattern and Flow Force
Corresponding to the Path of Hysteresis
Figure 15(a) and (b) show the axial flow force against flow rate change and valve displacement, respectively, inthe operating region where hysteresis occurs in the pressure-flow characteristic experiment as shown in Figure 3.
The axial flow force, flow rate, and valve displacement are non-dimensionalized by their respective values at the
rated flow rate. As in Figure 3, the red solid line indicates the forward path of hysteresis, and the blue solid line
indicates the return path of hysteresis. As for the axial flow force against the flow rate change shown in Figure
15, in the forward path, the negative flow force gradually increases as the flow rate starts to increase fromaround Q = 0 [ND], and unstable fluctuations start around the flow rate Q = 0.69 [ND], and then the flow force
suddenly becomes positive around Q = 0.83 [ND]. In the return path, as the flow rate begins to decrease fromaround Q = 1 [ND], the flow force decreases linearly and quickly drops to around 0 [ND] near the point where
the flow rate drops below Q = 0.42 [ND]. The positive and negative flow forces are of the same order.
The relationship among hysteresis, flow pattern, and axial flow force will be discussed. First, flow patterns 1, 2,
and 3 appear in the region below Q = 0.25 [ND].Then, as the flow rate increases, flow pattern 4 (Attached flow)
appears in the region where the negative flow force increases (Q = 0.3 ~ 0.69 [ND]). In this region (Q = 0.69 ~0.83 [ND]), both flow pattern 4 and flow pattern 5 appear in confusion, and the flow pattern is about to make atransition by sensing a slight disturbance. When the flow rate increases until it passes through this region, thestable flow pattern 5 (Straight flow) appears.
In short, we can conclude that severe oscillations occur between flow pattern 4 (Q = 0.3 ~ 0.69 [ND]) and flow
pattern 5 (Q > 0.83 [ND]) due to the transition phenomenon. This indicates that the hysteresis characteristic isdue to the transition of the flow pattern from 4 (Attached flow) to 5 (Straight flow).
In addition, the magnitude of the flow force becomes negative (-) and positive (+) in flow patterns 4 and 5,
respectively, which synchronizes the valve displacement with the change in flow force and thus the flow pattern.
Finally, the three-dimensional calculation results of the parameters ΔP, Q, and x, which are thought to govern
this phenomenon, are shown in Figure 16. There are two paths for the valve displacement, one that maintains thelinear relationship and the other that deviates from it, only within the region where the hysteresis characteristic
occurs. The flow pattern corresponding to this linear relationship is Straight flow, and the flow pattern deviatingfrom it is Attached flow. This result indicates that the Attached flow is caused by a small change in the valve
displacement.

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Axial force acting on the valve body
In case of valve closure
In case of valve release
Axial flow force
Figure 14. Calculation Way of Flow Force(a) Flow Force vs. Flow Rate(b) Flow Force vs. Valve Displacement
Figure 15. Time Series Data of Axial Flow Forces Corresponding to the Path of Hysteresis shown in Figure 3

Figure 16. Three-Dimensional Representation of Time Series Data of Pressure P, Flow Rate Q and Valve
Displacement x, which Correspond to the Path of Hysteresis shown in Figure 3

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CONCLUSION
In this study, CFD software was used to systematically organize the flow patterns in the prototype poppet valve
used in the study by Masuda et al. [2] and to clarify the relationship between hysteresis characteristics and flow
patterns.
The following points were clarified.
1. The flow patterns in the poppet valve can be classified into five categories.

2. In a poppet valve, a transition in flow pattern occurs between two stable flow patterns 4 and 5 (Attachedflow and Straight flow) as the flow rate increases. At that time, large pressure and valve displacement
oscillations occur.
3. Attached flow is caused by a small change in valve displacement.
4. As for the axial flow force, in flow pattern 4 (Attached flow), the negative flow force increases as the flowrate increases. On the other hand, in flow pattern 5 (Straight flow), the positive flow force increases as theflow rate increases. And the order of the positive and negative flow forces is about the same.
5. This difference in the polarity of the flow force affects the valve displacement and causes a transition in theflow pattern. The cause was found to be the collapse of the axial symmetry of the flow as a result of thecollapse of the axial symmetry of the circumferential pressure distribution in the separated bubble (lowpressure part) generated at the cylindrical corner of the poppet cone end as the flow rate increases.
In conclusion, the path of the hysteresis characteristics of the poppet valve is closely related to the flow patterns.

REFERENCES
1. OSHIMA, S., An Experimental Study on the Several Poppet Valves with Difference in Shape, Proceedings of the JFPSInternational Symposium on Fluid Power, 1989, 1989-1 pp. 361-368.
2. MASUDA, S., SHIMIZU, F., FUTIWAKI, M., TANAKA, K., Internal Flow and Hysteresis Characteristic of the Poppet
Type Pressure Control Valve, Transaction of the Japan Fluid Power System Society, 2021, 52-1, pp.8-15.
(MASUD, S., et al, Internal Flow and Hysteresis Characteristic of the Poppet Type Pressure Control Valve, This paper
has also been presented at JFPS International Symposium on Fluid Power hakodate2020.)
3. SCHRENK, E., Disc Valves, Flow Patterns, Resistance, and Loading, British Hydromechanics Research Association,
1957, 547, pp.1-17.
4. MCCLOY, D., MCGUIGAN, R. H., Some Static and Dynamic Characteristics of Poppet Valves, Proc. Instn. Mech.
Engrs, 1964, 179, pp199-213.
5. JOHNSTON, D. N., EDGE, K. A., Experimental Investigation of Flow and Force Characteristics of Hydraulic Poppet
and Disc Valves, Proc. Instn. Mech. Engrs, Part A, 1991, 205, pp161-171.
6. VAUGHAN, N. D., et. al., Numerical Simulation of Fluid Flow in Poppet Valves, Proc. Instn. Mech. Engrs, Part A,
1992, 206, pp.119-127.
7. HAYASHI, S., Instability of Poppet Valve Circuit, JSME International Journal. Ser. C, Dynamics, Control, Robotics,
Design and Manufacturing, 1995, 38-3, pp357-366.
8. ALTARE, G., RUNDO, M., OLIVETTI, M., 3D Dynamic Simulation of a Flow Force Compensated Pressure Relief
Valve, Proc. Instn. Mech. Engrs Congress and Exposition IMECE2016 November 11-17, IMECE2016-65624.
9. SIMERICS, Simerics Manual V5.0 (2019).

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Numerical Study on Identification Inputs for
Nonlinear Hydraulic Cylinders
Teruo KATO*, Satoru SAKAI*, Ryo ARAI**

Department of Mechanical Engineering, Shinshu University
4-17-1 Wakasato, Nagano City, Nagano, 380-0928 Japan(E-mail: 20hs203d@shinshu-u.ac.jp)

Abstract. This short report discusses the identification input for nonlinear hydraulic cylinders. First, we focus on an
input-output nonlinearity which exists even in the absence of nonlinear friction and prevents us from observing andmeasuring the input-output behaviors. Second, we propose a new identification input via a modification that isnumerically studied via a nondimensionalization technique. Finally, we confirm the effectiveness of the new
identification input via a numerical system identification.
Keywords: Hydraulic cylinder, Identification, Input-output nonlinearity.

INTRODUCTION
Hydraulic robots can achieve high power-weight ratio in comparison with electric robots and also realize agravity compensation without any control. A new modeling of hydraulic cylinder dynamics will be a foundation
to improve control performance against the nonlinearity, uncertainty, parameter perturbation, and so on (e.g., [1][2] [3]). In general, the modeling procedure (e.g., [4]) in a broad sense may include the identification procedure(e.g., a system identification procedure, a parameter identification procedure [5]). The identification procedure isalways needed in the numerical simulation procedure (e.g., [6]) at least in order not only to check the stability andcontrol performance before the experimental procedure and also to search a set of good parameters (e.g., link ratio,
control gain). Especially, in the context of learning (e.g., a deep learning) based on Big Data, the parameter
identification procedure is gaining importance.
However, in many cases, we are disturbed to observe and measure the input-output behaviors of nonlinear
hydraulic cylinders. It is well-known that the piston displacement cannot be small in the steady state even if thestandard identification input (e.g., the spool displacement for a servo-valve) such as a sinusoidal amplitude issmall [7]. For example, when we implement these sinusoidal signals for a nonlinear hydraulic cylinder, the piston
displacement in the steady state reaches and collides with one of the ends (the rod side) of the hydraulic cylinder.
This is because an input-output nonlinearity which exists even in the absence of nonlinear friction. Due to this
input-output nonlinearity, we cannot get enough observation and measurement to apply many system
identification and parameter identification methods. This is a serious problem in the identification procedure.

Figure 1. Hydraulic actuator.

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NOMINAL MODEL
Figure 1 shows a hydraulic arm and its cylinder. Let us review the conventional nominal model in the original
representation: [8] [9] [10] [11].
Consider the conventional nominal model{𝑀𝑠̈ = −𝐷𝑠̇ + 𝐴+ 𝑝+ − 𝐴− 𝑝−𝑏(−𝐴+ 𝑠̇ + 𝑄+ )𝑝̇+ =𝑉+ (𝑠)𝑏(+𝐴− 𝑠̇ − 𝑄− )𝑝̇− =𝑉− (𝑠)(1)

where the displacement 𝑠(𝑡) [m] , the cap pressure 𝑝+ (𝑡) [Pa] , the rod pressure 𝑝− (𝑡) [Pa] , and the spool
displacement (the input) 𝑢(𝑡) [m] are the functions of time 𝑡 [s]. The subscript + and − denote the cap-side andthe rod-side, respectively, and the subscript ± denotes both sides. The driving force is 𝑓(𝑡) = 𝐴+ 𝑝+ (𝑡) −𝐴− 𝑝− (𝑡) [N]. The mass 𝑀 [kg], the damping coefficient 𝐷 [Ns/m], the piston areas 𝐴+ ≥ 𝐴− [m2 ], and the bulk
modulus 𝑏 [Pa] are the positive constants. The cylinder volumes 𝑉+ (𝑠(𝑡)) = 𝐴+ (𝐿/2 + 𝑠(𝑡)) , 𝑉− (𝑠(𝑡)) =𝐴− (𝐿/2 − 𝑠(𝑡)) [m3 ] with the constant stroke 𝐿 [m] are the functions of the displacement 𝑠(𝑡)[m]. The input
𝑄+ and 𝑄− [m3 /s], are approximated by Bernoulli’s principle:
𝑄+ = 𝐶ℎ+ 𝑢, 𝑄− = 𝐶ℎ− 𝑢

(2)

with
ℎ+ ∶= √|sgn(−𝑢)𝑝+ +1 + sgn(+𝑢)𝑝𝑠 |2ℎ− ∶= √|sgn(+𝑢)𝑝− +1 + sgn(−𝑢)𝑝𝑠 |2where the flow gain 𝐶 [√m5 /kg] and the source pressure 𝑃 [Pa] are the positive constants. The conventional
nominal model introduces the restricted domain 𝑠 ∈ (−𝐿/2, 𝐿/2) and 𝑝± ∈ [0, 𝑝𝑠 ] and the absolute notation
within the square root functions (2) is dropped. The conventional nominal model (1) (2) in the original
representation is not our result but well-known. The equations (1) ignore the nonlinear friction effect and also theinternal and external leakage effects at least. Indeed, only the linear friction 𝐷𝑠̇ (𝑡) is taken into account in thenominal model. The equations (2) assume the steady flow and the negligible servo dynamics of the zero-lappedspool valve.
On the other hand, the stroke 𝐿 can include the pipe length effect and the bulk modulus 𝑏 includes the pipe
(or tube) flexibility effect. Of course, the difference (e.g., the nonlinear friction effect by the actual sealing)
between the conventional nominal model and the experimental setup exists and depends on each experimental
setup but would change continuously. In the context of robust control [12] [13], the difference is taken into accountin the controller design procedure in many controller design procedures (e.g. [14] [15] [16] [8] [9] [10]) as
uncertainty.
One of the linearized models of the nominal model is given as:{𝑀𝑠̂̈ = −𝐷𝑠̂̇ + 𝐴+ 𝑝̂+ − 𝐴− 𝑝̂−𝑏𝑝̂̇+ =(−𝐴+ 𝑠̂̇ + 𝐶√𝑝𝑠 /2 𝑢)
𝑉+ (0)𝑏𝑝̂̇− =(+𝐴− 𝑠̂̇ − 𝐶√𝑝𝑠 /2 𝑢)
𝑉− (0)(3)

whose state 𝑥̂(𝑡) ≔ (𝑠̂ (𝑡), 𝑠̂̇ (𝑡), 𝑝̂̇+ (𝑡), 𝑝̂̇− (𝑡))T starts from the same state 𝑥̂(0) = 𝑥(0) in the presence of thesame input 𝑢(𝑡).

INPUT-OUTPUT NONLINEARITY
We focus on an input-output nonlinearity which always exists even in the absence of nonlinear friction. Suppose
the sinusoidal signal:

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𝑢(𝑡) = 𝐴𝑢 sin(2𝜋𝑓𝑡) + 0(4)

is applied as the standard identification input. Even if the sinusoidal amplitude 𝐴𝑢 is small, it is well-known thatthe piston displacement 𝑠(𝑡) cannot be small in the steady state and reaches and collides with one of the ends (therod side) of the cylinder 𝑠(𝑡) → 𝐿/2 eventually [7]. Due to this input-output nonlinearity, we cannot get enoughobservation and measurement to apply many system identification and parameter identification methods. This isa serious problem in the identification procedure.
The following proposition is useful to understand this problem explicitly from a familiar linear approximation
point of view. Here, let us introduce the nondimensionalization in order to discuss comprehensively [17] [18] [19].
• Proposition 1 : Consider the original representation (1) (2) in the original representation. Then, there exists a
set of a time scaling 𝑡 ∗ = (1/𝑇)𝑡 , a variable scaling(𝑠 ∗ , 𝑝+∗ , 𝑝−∗ )T = ((1/𝑆)𝑠, (1/𝑃+ )𝑝+ , (1/𝑃− )𝑝− )T , and an
input scaling 𝑢 = (1/𝑈)𝑢 ∗ by which the standard sinusoidal signal:𝑢∗ (𝑡 ∗ ) = 𝐴∗𝑢 sin(2𝜋𝑓 ∗ 𝑡 ∗ ) + 0
causes the piston displacement in the steady state:𝑠 ∗ (𝑡 ∗ ) = 𝐵𝑢∗ sin(2𝜋𝑓 ∗ 𝑡 ∗ + 𝜙) + 0
where𝐵𝑢∗ = 𝐴∗𝑢 √8𝑃∗
𝜔 ∗6 + (𝐷 ∗2 − 4(𝐴∗ + 1))𝜔 ∗4 + 4(𝐴∗ + 1)2 𝜔 ∗2
𝜔∗2 − 2(𝐴∗ + 1)
𝜙 = tan−1 ()−𝐷 ∗ 𝜔 ∗with 𝜔∗ = 2𝜋𝑓 ∗Let us recall the set of the time transformation 𝑡 ∗ = (1/𝑇)𝑡 with 𝑇 = √𝑀𝐿/(𝑏𝐴), the state transformation∗𝑥 = ( 𝑠 ∗ , 𝑠̇ ∗ , 𝑝+∗ , 𝑝−∗ )T = 𝑋𝑠−1 𝑥 with𝑆
0
𝑋𝑠 ≔ [
0
0

0
𝑀𝑆/𝑇
0
0

0
0
𝑃+
0

𝐿
0
0
0
]=[
0
0
𝑃−
0

0
√𝑀𝐿𝑏𝐴+
0
0

0
0
𝑏
0

0
0
],
0
𝑏

and the input transformation 𝑢∗ = (1/𝑈)𝑢 with 𝑈 = (√𝐴3+ 𝐿/𝑀/𝐶). The original representation (1) (2) in theoriginal representation is transformed to the following nondimensional representation:{𝑠 ∗′ = −𝐷 ∗ 𝑠 ∗′ + 𝑝+∗ − 𝐴∗ 𝑝−∗1(−𝑠 ∗′ + 𝑄+∗ )𝑝+∗′ =
1/2 + 𝑠 ∗1(+𝑠 ∗′ − (1/𝐴∗ )𝑄−∗ )𝑝−∗′ =
1/2 − 𝑠 ∗and𝑄+∗ = 𝐶ℎ+∗ 𝑢, 𝑄−∗ = 𝐶ℎ+∗ 𝑢
withℎ+∗ ∶= √|sgn(−𝑢)𝑝+∗ +1 + sgn(+𝑢) ∗𝑃 |2ℎ−∗ ∶= √|sgn(−𝑢)𝑝−∗ +1 + sgn(−𝑢) ∗
𝑃 |
2

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where 𝑇, 𝑆, 𝑃+ , 𝑃− , 𝑈 are the constants and 𝐷 ∗ , 𝐴∗ , 𝑃 ∗ are the nondimensional parameters defined as 𝐷 ∗ ≔𝐷√𝐿/(𝑀𝑏𝐴+ ), 𝐴∗ ≔ 𝐴− /𝐴+ (≤ 1), 𝑃 ∗ ≔ 𝑝𝑠 /𝑏. The notation ●′ denotes the derivative with respect to thenondimensional time 𝑡 ∗ . Simultaneously, the linearized models (3) after the standard linear approximation has
the following transfer function matrix the nondimensional transfer function matrix:(+𝑝2 + 𝐷 ∗ 𝑝 − 2(1 − 𝐴∗ ))/(2𝑝)√8𝑃(−𝑝2 − 𝐷 ∗ 𝑝 − 2(1 − 𝐴∗ ))/(2𝐴∗ 𝑝)2∗∗
𝑝 + 𝐷 𝑝 + 2(1 + 𝐴 )𝑝 + 𝐷∗
[
1/𝑝
]
∗

(5)

from the input 𝑢∗ to the estimated outputs 𝑦̂0∗ = (𝑝̂+∗ , 𝑝̂−∗ , 𝑓̂ ∗ , 𝑠̂ ∗ )T of the linearized model (3) in thenondimensional version. Here, the notation 𝑠 denotes the derivative operator in the Laplace transform withrespect to the nondimensional time 𝑡 ∗ = (1/𝑇)𝑡. This immediately brings the explicit formula on the gain andphase. Note that the above proposition (5) uses the similar technique in our paper [19] but the explicit formula on
the gain and phase has never reported. Note that the obtained formula:√8𝑃∗
𝑝3 + 𝐷 ∗ 𝑝2 + 2(1 + 𝐴∗ )𝑝with the restricted parameter 2(1 + 𝐴∗ ) ≤ 4 is different from the familiar formula:𝐾1

𝑝2 + 2𝜁𝑤𝑛 𝑝 + 𝑤𝑛 2 𝑝
with the unrestricted parameter 𝑤𝑛 2 < ∞. This difference implies a link to the conservative control performance[13] based on the familiar formula. In a word, the familiar formula is too large and is not used here in this paper.
Based on Proposition 1, we can explicitly design a standard identification input with the amplitude 𝐴∗𝑢 to achievethat the piston displacement with the amplitude 𝐵𝑢∗ does not reach and collide with one of the ends (the rod side)of the cylinder comprehensively. However, this is for the linearized model. We should design a certain
identification input to achieve for nominal model of nonlinear hydraulic arms. This is an understanding of theserious problem from a familiar linear approximation point of view.

NEW IDENTIFICATION INPUT
As the main contribution of this paper, in this section, we propose a new identification input modification for
nonlinear hydraulic cylinders.
• Condition 1 : The new identification input should be defined not in the feedback manner but the feedforwardmanner. In general, in identification procedure (in modeling procedure in a broad sense), it is of importance toobserve and measure the input-output behavior as it is. However, in many cases, control inputs constructing someclosed-loop systems change the input-output behaviors even when the linear approximation is not needed. We
should define the identification input only use a feed-forward manner.
• Condition 2 : The new identification input should close to the standard identification input (e.g., the standard
sinusoidal input) in some senses. This is because, in many cases, we already have a few knowledge of the system
and we cannot confirm the safety of the system. In other words, we should avoid a large modification from thestandard identification input.
As one of the simplest new identification inputs, we propose the following modification:𝑢∗ (𝑡 ∗ ) = 𝐴∗𝑢 sin(2𝜋𝑓 ∗ 𝑡 ∗ ) + 𝑢0∗ (𝐷 ∗ , 𝐴∗ , 𝑃∗ , 𝑓 ∗ )(6)where 𝑢0∗ is the nondimensional constant. The first term is nothing but the standard sinusoidal input (4), and thesecond term is a special constant depending on the three nondimensional parameters as well as the frequency
appeared in the first term. Clearly, the proposed identification input (6) is given in a feed-forward manner and theonly difference between them is just the existence of the constant 𝑢0∗ . The above 2 conditions (Condition 1 andCondition 2) are satisfied in these senses. The remaining problem is how to search the constant 𝑢0∗ . Now let us
define the best constant 𝑢0∗ as a minimization of 𝑢0∗ to achieve the following formula:

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𝐽(𝑢0∗ ) =𝑠 ∗ (0) − min0≤𝑡 ∗≤𝑇𝑢∗ 𝑠 ∗ (𝑡 ∗ )0.01 (max0≤𝑡 ∗≤𝑇𝑢∗ 𝑠 ∗ (𝑡 ∗ ) − min0≤𝑡 ∗≤𝑇𝑢∗ 𝑠 ∗ (𝑡 ∗ ))> 1

as an objective function in this paper. Here 𝑇𝑢∗ is the test time. Since the search space of the constant 𝑢0∗ is justthe 1 dimensional space, we apply an advanced direct search method in which we repeat the nonlinear dynamics
computations and the parameter updates in the only 3 dimensional parameter space (𝜃 ∗ ≔ (𝐷 ∗ , 𝐴∗ , 𝑃 ∗ ) ∈ℝ × [0,1] × ℝ ⊂ ℝ3 ). Unfortunately, many experimental conditions are often restrictive in terms of the scale andcannot be comprehensive. On the other hand, the usual numerical conditions are also restrictive when theparameter space is too large. However, the advanced direct search method is quite efficient than the conventional
direct search method with the 8 dimensional parameter space(𝜃 ≔ (𝑀, 𝐷, 𝐿, 𝐴+ , 𝐴− , 𝑏, 𝐶, 𝑝𝑠 ) ∈ ℝ8 ). Due tothe parameter reduction (8 to 3), the computational time complexity of the advanced direct search method has
𝒪 (𝑛3 ) and that of the conventional one has 𝒪 (𝑛8 )(we see 𝑛 = 10 case in the next section). Furthermore, insteadof the original state, the Casimir functions reduce the computational time [20]. In the nonlinear dynamics
computations, the Casimir functions are applied to compute the non-dimensional state 𝑥 ∗ (𝑡 ∗ ) starting from theinitial state 𝑥 ∗ (0) = (0, 0, 𝑃 ∗ /2, 𝐴∗ 𝑃 ∗ /2)𝑇 in the presence of the new identification input𝐴∗𝑢 sin(2𝜋𝑓 ∗ 𝑡 ∗ ) + 𝑢0∗ with the amplitude 𝐴∗𝑢 ≔ 𝐴𝑢 /𝑈𝑠 and the frequency 𝑓𝑢∗ ≔ 𝑇𝑠 𝑓𝑢 ∈ {0.002, 0.15}. The test
period is defined as [0, 𝑇𝑢∗ ]:=[0, 3/𝑓𝑢∗ ] and the Adams-Moulton method with the variable step is applied(MATLAB 2015a, Simulink ver.8.6, 64 [bit], CPU 1.60 [GHz], Memory 8.0 [GB]).

Figure 2. The best constant 𝑢0∗ in 𝐷 ∗ 𝐴∗ 𝑃∗ -space (𝐷 ∗ = 𝐷√𝐿/(𝑀𝑏𝐴+ ), 𝐴∗ = 𝐴− /𝐴+ , 𝑃 ∗ = 𝑝𝑠 /𝑏).

Figure 3. The value of 𝑢0∗ 𝐴∗2 in 𝐷 ∗ 𝐴∗ 𝑃 ∗ -space (𝐷 ∗ = 𝐷√𝐿/(𝑀𝑏𝐴+ ), 𝐴∗ = 𝐴− /𝐴+ , 𝑃∗ = 𝑝𝑠 /𝑏).
Figure 2 shows a part of the searched constant 𝑢0∗ in the 3 parameter-space for each frequency. As an
unexpected result, the constant 𝑢0∗ mainly depends on the nondimensional area 𝐴∗ only and less sensitive to theother nondimensional physical parameters. Fig. 3 shows the value of 𝑢0∗ 𝐴∗2 in the 3 dimensional parameter
space for each frequency. One of the simplest approximation is numerically obtained as follows:𝑢0∗ ≈ 𝑢0∗ (𝐴∗ ) ≔ −0.0266/𝐴∗ 2(7)

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EXPERIMENTS
Figure 4 shows an appearance of the experimental setup and Fig. 5 shows the signal diagram. The experimental
setup consists of a manipulator (the arm length 315 [mm]), a pump (DAIKIN, NDR081-071H-30, 11.7 [L/min],𝑝𝑠 = 7 [MPa]), a tank 7 [L], four pipes (YOKOHAMA, SWP70-6, 1/4 [inch]), a valve (YUKEN KOGYO, LSVG01EH-20-WC-A1-10), two cylinders asymmetric cylinder (JPN, KW-1CA30×75, L=75 [mm]), a filter (Taisei
Kogyo, UM-03-20U-1V) and an oil (ISOVG32, 860 [kg/m3 ], 40 ± 2 [℃]).
The output signals 𝑝± are measured by a pressure sensor (KEYENCE, AP-15S), the output 𝑠 is measured bya potentiometer (MIDORI, LP-100F-C) and the identification input 𝑢 is measured by the LVDT. These output
and input signals are detected by an AD converter (Interface, PCI-3155, 16 [bit]) and recorded by a control PC(EPSON, LX7700, Linux [21], 2.53 [GHz]) by a sampling time 1.0 [ms]. In order to identify the parameters, achirp-like input with the amplitude 𝐴𝑢 = 1.2 [V], the frequencies 𝑓 = 0.5 + (5.0 − 0.5)/36 ∙ 𝑡 [Hz] is applied
via a DA converter (Interface, PCI-3325, 12 [bit]). The identified and measured parameters are(𝑀, 𝐷, 𝐿, 𝐴+ , 𝐴− , 𝑏, 𝐶, 𝑝𝑠 ) = (14, 3200, 0.075, 7.0 × 10−4 , 5.4 × 10−4 , 5.3 × 108 , 1.6 × 10−4 , 7.0 ×
106 )[5]. Since the parameter identification is in the off-line world, the time derivative of the piston displacement𝑠 and the pressures 𝑝± are sufficiently calculated via the first or second difference approximation and thestandard moving average.

Figure 4. The experimental setup (photo).

Figure 5. The experimental setup (signal flow).

Figure 6. Experimental setup (Black) vs. Nominal model (Red) (only piston velocity of Nominal model is
not Red but White)[11].

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Figure 6 shows a long time cross validation in which the experimental outputs (Black) were never used in theparameter identification procedure. Nevertheless, with respect to the nonlinearly-shaped responses in the cappressure, the nominal model has an accuracy that any linearized model (transfer function) cannot have. Fig. 7 andFig. 8 are simulation results and experimental results. Those results show the time response of the piston
displacement in the presence of the conventional chirp-like input and also shows the time response of the piston
displacement in the presence of the same chirp-like input with the proposed constant 𝑢0∗ in the presence of same
initial states. Even though the applied constant 𝑢0∗ is 0.046 and very tiny (about 4.6 [%] compared with the input
sinusoidal amplitude 𝐴∗𝑢 ), the input-output nonlinearity is almost disappeared. Eventually, the proposed constant𝑢0∗ avoids the collision with the end of the cylinder 𝑠(𝑡) ↛ 𝐿/2. Now we can fully observe and measure theinput-output behaviors. On the other hand, in case of the conventional chirplike input without the constant 𝑢0∗ ,not only the identification procedure but also the model validation procedure can fail. The second impact of the
proposed constant 𝑢0∗ is small. This means that our approach using the new identification input (7) is very
effective to solve the problem.

Figure 7. Simulation results.

Figure 8. Experimental results.

NUMERICAL SYSTEM IDENTIFICATION
Figure 9 shows the bode diagram of the nominal model. Because of new identification input, we can fully
observe and measure the input-output behaviors to apply a numerical system identification, numerically.

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Figure 9. The identified bode diagram of the nominal model. X markers are numerical outputs (gain of inputput) which are observed.

CONCLUSION
The report provides a very simple but comprehensive modification of the standard identification input for
nonlinear hydraulic cylinders. Due to the input-output nonlinearity, we cannot observe and measure the inputoutput behaviors to apply many system identification and parameter identification methods. This is a serious
problem in the identification procedure. We propose a new identification input via the modification consisting ofa very simple but comprehensive constant 𝑢0∗ . We show the effectiveness of the proposed identification input.
The piston displacement can be small even in the steady state and we can fully observe and measure the inputoutput behaviors and can success the numerical system identification.

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12. A. J. van der Schaft. L2-Gain and Passivity Techniques in Nonlinear Control. Originally published as volume 218 in theseries: Lecture Notes in Control. Springer-Verlag London, London, 2000.
13. K. Zhou, J. C. Doyle, and K. Glover. Robust and Optimal Control. Prentice-Hall, Inc., Upper Saddle River, N.J., 1996.
14. B. Yao, F. Bu, J. Reedy, and G. T.-C. Chiu. Adaptive robust motion control of single-rod hydraulic actuators: Theory
and experiment. IEEE/ASME Transactions of Mechatoronics, 2000, 5, pp. 79–91.
15. T. Boaventura, M. Focchi, M. Frigerio, J. Buchli, C. Semini, G. A. Medrano-Cerda, and D. G. Caldwell. On the role ofload motion compensation in high-performance force control. In Proc. IEEE/RSJ IROS, 2012, pp. 4066-4071.
16. Jouni Mattila, Janne Koivumaki, Darwin G. Caldwell, and Claudio Semini. A survey on control of hydraulic robotic
manipulators with projection to future trends. IEEE/ASME Transactions on Mechatoronics, 2017, 22(2) pp. 669-680.
17. I. Nezu and H. Nakagawa. Turbulence in Open Channel Flows. CRC Press, 1993.
18. W. D. Curtis, J. David Logan, and W. A. Parker. Dimensional analysis and the pi theorem. Linear Algebra and Its
Applications, 1982, (47) pp. 117-126.
19. E. Buckingham. On physically similar systems. Physics review, 1931, 4, pp. 354-376.
20. S. Sakai. Casimir based fast computation for hydraulic robot optimizations. In Proc. IEEE/RSJ IROS, 2013, pp. 28742881.
21. S. Sakai, K. Osuka, T. Maekawa, and M. Umeda. Robust control systems of a heavy material handling agricultural robot:
A case study for initial cost problem. IEEE Transactions on Control Systems Technology, 2007, 15, pp. 1038-1048.

8

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On the Analysis of Energy Behaviors in Hydraulic Cylinder
Dynamics via Modeling of Experimental Excavators
Ryo ARAI *, Satoru SAKAI*, Akihiro TATSUOKA***

Department of Mechanical Engineering, Shinshu University
4-17-1 Wakasato, Nagano city, Nagano, 390-8553 Japan(E-mail: 20w4001g@shinshu-u.ac.jp, satorus@shinshu-u.ac.jp)**
Mitsubishi Heavy Industries

Abstract. The paper partially discusses an analysis of the nominal model of hydraulic cylinder dynamics. The pump pressure
dynamics is especially focused. First, we review hydraulic cylinder dynamics and apply our physical parameter identification
method to agriculture scale experimental excavators. Second, we analyze the energy behaviors by one of the Hamiltonians thatis different from the previous Hamiltonian with the exponential terms.
Keywords: Hydraulic Machinery, Physical Parameter Identification, Energy, Port-Hamiltonian Theory

INTRODUCTION
Unmanned hydraulic machineries are gaining popularity and a new generation of model based control of hydraulicmachinery is coming (e.g., [1][2][3][4][5]). In this paper, we discuss energy behaviors of hydraulic cylinder
dynamics which is important for model based control of agriculture scale excavators. In agriculture scale, the mostwidely used controls are PID or PID-like control. However, model based control has the possibility to achieve
better performance than PID control [6][7][8]. Especially, even against modeling errors, the robust control theory
such as the port-Hamiltonian theory guarantees robustness [9][10]. In the paper, in agriculture scale, we discuss
the experimental, numerical, and analytical energy behaviors of a Hamiltonian in hydraulic cylinder dynamics.
The paper has two main results. As the first main result, our previous physical parameter identification method[11][12] is applied to experimental excavators that have a pump pressure dynamics. As the second main result,
we discuss the experimental, numerical, and analytical energy behaviors in hydraulic cylinder dynamics from theport-Hamiltonian point of view by one of the Hamiltonians that is different from the previous Hamiltonian withthe exponential terms [13].

NONLINEAR NOMINAL MODEL
Let us consider a hydraulic arm whose nonlinear nominal model is a coupling of the hydraulic cylinder dynamics[14] and a pump pressure dynamics defined later. First, the hydraulic cylinder dynamics is described by

(1)
with
where the subject 𝑖 = 1 or 2 is the joint number and the subject + or − is denotes the cap side and the rod side.
The joint angle[rad], the joint angle velocity[rad/s], the cap side pressure[Pa], the rod side[m2 ] are the functions of time 𝑡 [s]. The drift term 𝑓0 (𝑥) and the input
pressure[Pa], and the orifice areaterm
are the functions of the state 𝑥. Second, the pump pressure dynamics is described by(2)[m3 /s] are the function of time. The pump bulk
where the pump pressure[Pa] and the pump flow rate3
[m3 /s] and themodulus 𝑏𝑝 [Pa] and the pipe volume 𝑉𝑝 [m ] are the physical parameter. The input flow rate3
[m /s] are the function of the tank pressuretank flow rate
[Pa] and approximated by Bernoulli's principle
as hydraulic cylinder dynamics.

1

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PHYSICAL PARAMETER IDENTIFICATION AND VALIDATIONIn this section, we update our physical parameter identification method [11][12] to agriculture scale excavators.
Suppose thatis obtained at each time
From theproject theorem in Hilbert space [15], the set of the unknown physical parameterandare identified uniquely
as the following values:(3)
where the matrices 𝑋1𝑁 , 𝑌1𝑁 , 𝑋2𝑁 , and 𝑌2𝑁 are obtained from the experimental outputs.
The experimental setup consists of an agriculture scale excavator (ZX30U-5A), pipes (PFH06, 3/8 inch), and a
servo open-center valve only. The excavator is driven by the servo open-center valve whose the orifice area
𝑢2 (𝑡) is measured by a displacement sensor (LP-20F) and a DC servo motor (T720-0.12EL8), a rack and a pinion(reduction ratio 1/25), an AD board (PCI-3155, 16 bit), a control PC (Linux, 3.8 GHz, 4.0 GB), and a DA board(PCI-3325, 12 bit). The outputs of the excavator are a piston displacement measured by a piston displacement
sensor (CPP-45-150LS), the cap, rod, pump, and tank pressures measured by the pressure sensors (AP-16S). Since
the physical parameter identification is in the off-line case, the time derivative of the piston displacement
𝑠̇2 (𝑡) and the cap and rod sides pressures
are sufficiently estimated via the first difference approximation.
[m2 ] is applied in
In order to identify the physical parameters, the ideal sinusoidal signal
a period (𝑡 = 2.0 − 20.0) [s].
Table 1.

Identification result for the nominal numerical simulator.

Table 1 shows the results of the identification. All the identified physical parameter values were positive without
any restrictions on the sign or value range of the parameters
This section also discusses the validation. First, the nominal numerical simulator is constructed with the physicalparameters on the laptop (2.2 GHz, 4.0 GB, MATLAB R2015a, Simulink). Second, by the Adams-Bashforthmethod (ode113) the numerical outputs are generated by an experimental input 𝑢2 (𝑡). Third, the experimental
outputs and the corresponding numerical outputs are generated by the same experimental input and are comparedwith each other. The frequency of the ideal sinusoidal signal for validation is 𝑓 = 0.5 [Hz]. Here the initial statevalues of the nominal numerical simulator is given from the measured values except the initial piston
velocity 𝑠̇2 (𝑡) that is not measured but estimated by the just first order differentiation.

Figure 1.

Input and the outputs (Validation, f =0.5 [Hz], Joint 2).

2

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Figure 1 shows the results of the validation. In Figure 1, against modeling errors, not only the cap and rod sides
pressures
andbut also the pump pressure
are generated sufficiently. The amplitude error of thepump pressure can be observed but the phase and the number of the peaks were very similar.

ENERGY BEHAVIOR ANALYSIS
In this section, we compare the experimental, numerical, and analytical energy behaviors of the excavator in
agriculture scale. One of the most famous representations is the following input-state-output representation [9][10](4)where 𝑧 ∈ 𝐑𝑛 is the state, 𝑢̅ ∈ 𝐑𝑚 is the input and 𝑦̅ ∈ 𝐑𝑚 is the corresponding output. 𝐽(𝑧) = −𝐽T (𝑧) ∈𝐑𝑛×𝑛 is a skew symmetric matrix and 𝑅(𝑧) = −𝑅T (𝑧) ∈ 𝐑𝑛×𝑛 is a symmetric matrix. The notation ∇𝑧 is thepartial derivative with respect to 𝑧. In case of hydraulic cylinder dynamics, 𝐻(𝑧) are given as:(5)which is one of the Hamiltonians and different from the previous Hamiltonian with the exponential terms [13].
For the Hamiltonian in the paper, the first term is the kinetic energy. The second and third terms are the internalenergy. The fourth term is the potential energy. The cap and rod side pipe volumes 𝑉±2 (𝑠2 ) are the functions of
the piston displacement 𝑠2 (𝑡). The moment of inertia 𝐽2 , the gravity coefficient 𝑊2 , the joint angle 𝑞1𝑐𝑜𝑛𝑠𝑡 , andthe gravity acceleration , and the bulk modulus 𝑏 are the physical parameters. The time derivative of the
Hamiltonian
from the port-Hamiltonian point of view is given as:(6)
Now we can compare the experimental, numerical, and analytical time responses of the time derivative of the
Hamiltonian 𝐻̇ (𝑧) of the hydraulic cylinder dynamics with the identified physical parameters. The experimental
time responses of the time derivative of the Hamiltonian is derived directly from the measured Hamiltonian andits derivative, while the numerical and analytical time responses of the time derivative of the Hamiltonian are
derived indirectly by Equation (6). In the paper, we analyze the two cases, that is, the slowest case 𝑓 =0.5 [Hz] and the fastest case 𝑓 = 1.0 [Hz].

Figure 2.

Experimental (black) vs. Numerical (red) vs. Analytical (green) time responses of the time derivative of the
Hamiltonian. (Energy behavior analysis, Slow: 𝑓 = 0.5 [Hz], Fast: 𝑓 = 1.0 [Hz]).

Figure 2 shows the experimental, numerical, and analytical time responses of the time derivative of the
Hamiltonian in the slow case and fast case. The experimental, numerical, and analytical energy behaviors wereclose to each other. One of the most important features is the asymmetric behavior of the time derivative
Hamiltonian even though input is symmetric in the sense of the standard sinusoidal signal. This asymmetricalbehaviors comes from drift of the piston displacement 𝑠2 (𝑡) in Figure 1. This result implies that the portHamiltonian point of view will be applicable in agriculture scale against modeling errors.

3

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CONCLUSIONSIn this paper, a method for identifying physical parameters of an open-centered multi-degree-of-freedom hydraulicarm is proposed, and the effectiveness of the proposed method is verified by model validation and energy analysis
for a 1-DOF real construction machine arm. By calculating the output of the Hamiltonian, the validity of updating
the model to the Port-Hamiltonian system of an open-centered hydraulic arm was evaluated. In all, via theexperimental, numerical, and analytical energy behaviors, the applicability of the port-Hamiltonian theory is
confirmed in agriculture scale against modeling errors. Future work will be to verify the model and withsimultaneous motion of two axes, and to analyze the energy.

ACKNOWLEDGMENTS
This work was partially funded by JSPS KAKENHI Grant Number 21K03934. This funding is highly appreciated.
The authors are also grateful for the technical help provided by Yusuke Nabana, Yuki Nishimura, and Kazuki
Nagai.

REFERENCES
1. Zhang, Q., Basics of Hydraulic Systems, Taylor & Francis, 2008.
2. Huang, Y., Zhang, Q., Agricultural Cybernetics, Springer. 2021.
3. Sakai, S., Iida, M., Osuka, K., Umeda, M., Design and control of a heavy material handling manipulator for agriculturalrobots, Autonomous Robots, Vol. 25, No. 3, 2008, pp. 189-204.
4. Mattila, J., Koivumaki, D., Caldwell, D. G., Semini, C., A survey on control of hydraulic robotic manipulators withprojection to future trends, IEEE/ASME Trans. Mechatronics, Vol. 22, No. 2, 2017, pp. 669-680.
5. Raibert, M., Blankespoor, K., Nelson, G., Playter, R., BigDog, the rough-terrain quadruped robot, In Proceedings of the
IFAC World Congress, 2008, pp. 10822-10825.
6. Merritt, H., Hydraulic Control Systems, John Willey & Sons, 1967.
7. Watton, J., Fluid Power Systems, Englewood Cliffs, New Jersey, USA, 1989.
8. Mohieddine, J., Andreas, K., Hydraulic Servo-Systems, Springer, 2003.
9. Duindam, V., Macchelli, A., Stramigioli, S., Bruyninckx, H., Modeling and Control of Complex Physical Systems: The
Port-Hamiltonian Approach, Springer, 2009.
10. van der Schaft, A. J., L2-gain and passivity techniques in nonlinear control, Springer-Verlag, 2000.
11. Maeshima, Y., Sakai, S., Nakanishi, M., Osuga, K., A Base Parameter Identification Method and Model Validation for
Hydraulic Arms, Journal of JFPS, Vol. 43, No. 1, 2012, pp. 16-21.
12. Sakai, S., Maeshima, Y., A New Method for Parameter Identification for N-DOF Hydraulic Robot, In Proceedings of IEEE
International Conference on Robotics and Automation (ICRA), Hong Kong, China, 2014, pp. 5983-5989.
13. Arai, R., Sakai, S., Tatsuoka, A., Zhang, Q., Analytical, Experimental, and Numerical Investigation of Energy in Hydraulic
Cylinder Dynamics of Agriculture Scale Excavators, Energies, submitted.
14. Sakai, S., Stramigioli, S., Visualization of Hydraulic Cylinder Dynamics by a Structure Preserving
Nondimensionalization, IEEE/ASME International Transactions on Mechatronics, Vol. 23, No.5, 2018, pp. 2196-2206.
15. Luenberger, D., Optimization by vector space method, Wiley interscience, 1968.

4

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Organized Session | Organized session

Aqua Drive 1[OS4-1-01] Consideration of Thermal Stability of the Ultra-Precision WaterLubricated Spindle〇Dmytro Fedorynenko1, Yohichi Nakao2 （1.Tohoku University, 2.Kanagawa University）[OS4-1-02] Multidisciplinary Design Optimization of a Tortuous Path Trim for a
Labyrinth Control Valve〇Runlin Gan1, Xukang Li1, Song Liu1, Baoren Li1 （1.Huazhong University of Science&Technology, FESTO Pneumatic Technology Center）[OS4-1-03] Design of Magnetostrictive Power Generation Device from Pulsating
Pressure in Hydraulic Pipeline by Using Water Hydraulic Cylinder〇Kaito Miyashita1, Shouichiro IIO1, Tsuyoki TAYAMA2, Ryuichi ONODERA2, Shyota ABE2
（1.Shinshu University, 2.Tohoku Steel Co., Ltd.）[OS4-1-04] Flow Characteristics of a Cavitating Jet through a Small RectangularOrifice with Different Aspect Ratios〇Hironori Takei1, Kohei Terakawa1, Shouichiro Iio1, Kotaro Takamure2, Tomomi Uchiyama2,
Futoshi Yoshida3 （1.Shinshu University, 2.Nagoya University, 3.KYB Corporation）©The Japan Fluid Power System SocietyOS4-1-01

The 11th JFPS International Symposium on Fluid Power 2020

Consideration of Thermal Stability of the Ultra-Precision WaterLubricated Spindle
Dmytro FEDORYNENKO*, Yohichi NAKAO*** Department of Aerospace Engineering, Tohoku University
6-6-01 Aramaki-Aza-Aoba, Aoba-ku, Sendai, 980-8570 Japan(E-mail: fedorynenko.dmytro.b3@tohoku.ac.jp)** Department of Mechanical Engineering, Kanagawa University
3-27-1 Rokkakubashi, Kanagawa-ku, Yokohama 221-8686, Japan(E-mail: nakao@kanagawa-u.ac.jp)
Abstract. In this study, the thermal stability of the spindle with water-lubricated hydrostatic bearings was investigated
via simulation. The simulation study was divided into two separate tasks. As the first stage, the temperature distribution
in the flowing part of spindle bearings was determined utilizing the SolidWorks Flow Simulation. The experimentallygathered temperature of lubricant and spindle surfaces was used as boundary conditions for a finite volume problem.
Then, the thermal analysis of the spindle was performed using the obtained temperature field in the first stage. The
pattern of gap variation in journal and thrust bearings due to the thermal deformation of bearing surfaces wasestablished. The thermal stability of a spindle under the influence of an external temperature field is analyzed using abearing film force and stiffness as indicators.
Keywords: water-lubricated bearing, thermal deformation.

INTRODUCTION
Thermal deformation of ultra-precision spindles is the main limiting factor for improving machining accuracy andproductivity. The primary sources of heat in machine tool spindles are bearings. Aerostatic and hydrostatic spindle
bearings are widely used for ultra-precision machine tools [1]. High thermal stability at high speeds is the main
benefit of aerostatic bearings. However, they have very low damping and stiffness characteristics compared to oilbearings. Oil hydrostatic bearings are advantageous for ultra-precision spindles in terms of achieving highaccuracy of machining. When lubricating oil hydrostatic bearings, power losses due to oil flow cause heatgeneration, especially at high speeds [2]. The pressure power losses are caused by the pressure drop in the bearing
supply system and cannot be avoided. The power losses on viscous friction, especially at high speeds, lead to thetemperature rise of lubricating fluid and, consequently, to spindle deformation. Accordingly, it decreases
machining accuracy.
In contrast to oil, water has relatively low viscosity, higher specific heat, and thermal conductivity. Thus, it isconsidered that higher thermal stability can be achieved by using water as lubricating fluid for ultra-precisionspindles. A spindle with water-lubricated hydrostatic bearings was proposed to cope with the disadvantage of boththe aerostatic and oil hydrostatic bearings [3]. Static characteristics and dynamic behavior of the water-lubricatedspindle were then studied [4]. The patterns of the change in a spindle temperature field and the thermal stabilityof spindles with different rated rotational speeds were experimentally investigated [5]. However, the influence ofthermal deformations on the water-lubricated spindle accuracy has not been thoroughly investigated yet.
This paper aims to investigate the influence of thermal deformations on the thermal stability of the waterlubricated spindle. The simulation studies reveal the influence of the thermal deformations on the linear expansionof the spindle bearing surfaces, the gap change in bearings, and their stiffness.

STRUCTURE OF THE WATER-LUBRICATED SPINDLE
The investigated ultra-precision spindle with water-lubricated hydrostatic bearings [5] is schematically presented
in Fig. 1. The spindle has four recessed journal bearings and four recessed opposed-pad thrust bearings with shortpipe restrictors for each recess. The mentioned restrictors are designed and installed to achieve the highest static
stiffness for water-lubricated bearings [5]. The thrust bearing has an outer diameter of 89 mm and an innerdiameter of 51 mm. The nominal gap size in the thrust bearing is 10 μm. Journal bearings are located at both endsof the rotor. The journal diameter and the journal bearing's axial length are 50 mm and 49 mm, respectively. The
nominal gap size in the journal bearing is 12 μm.
The spindle rotor is made of stainless steel. Diamond-like carbon is plated onto the outer surfaces of the rotor andthe inner surfaces of the casing to prevent undesirable contact between them during spindle rotation [5]. The rated
supply pressures are 2.5 MPa and 2.1 MPa for journal and thrust bearings, respectively. The highest rotational
speed of the spindle is 3000 min−1.

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Figure 1. Structural Diagram of Water-Lubricated Hydrostatic Spindle

LUBRICATION TEMPERATURE FIELD
Temperature distribution and the resultant deformation of the spindle are keys of importance to develop an
effective temperature control technique aimed at achieving higher thermal stability of the ultra-precision spindle.
Temperature changes of the water flow and spindle have been evaluated via calculations and experiments. A
commercially available CFD is used for solving a heat transfer task in the spindle with internal fluid flow. The
temperature distribution inside a spindle-flowing part was derived by numerical simulation using experimentallyobtained data as boundary conditions to enhance the validity and accuracy of numerical analysis.
The simulations were carried out in two stages. At the first stage, the temperature field of lubrication in the spindle
was determined using the SolidWorks Flow Simulation, including heat transfer to a solid body. The data obtained
experimentally were used as the initial and boundary conditions: inlet water temperature, supply pressure, andoutlet flowrate (Tab. 1). As the outer wall temperature and initial solid temperature, the experimentally obtainedouter casing temperature and outer rotor temperature were taken, respectively. Finally, experimentally gatheredvalues of the outlet water temperature were used to evaluate the accuracy of numerical simulation.
Table 1. Boundary Conditions

Inlet Parameters
Supply water temperature, [°C]Supply pressure (journal bearing), [MPa]Supply pressure (thrust bearing), [MPa]Outlet Parameters
Total water flowrate, [l/min]
Additional Parameters
Outer wall temperature, [°C]
Initial solid temperature, [°C]Surface roughness, [µm]

23.72
2.5
2.1
4.2
23.93
22.88
0.8

Figure 2 presents the water temperature distribution on the bearing surfaces of a left journal bearing and the left
side of an opposed-pad thrust bearing, as it is depicted in a cross-section view in Fig. 1. High fluid temperature isformed on bearing lands achieving the maximum value of 24.9 °C on inter-recess lands of the thrust bearing. Thisphenomenon is due to a relatively small gap in the thrust bearing resulting in temperature rise when liquid flows
under pressure through narrow slots between recesses. On the contrary, the minimum temperature of lubricant isobserved in such areas of a bearing as recesses (24.33 °C) and drainage channels (24.21 °C) due to their relatively
high depth comparable to lands.

THERMAL ANALYSIS OF SPINDLE
At the second stage, the thermal analysis of the spindle was performed employing the SolidWorks Simulation.
The mentioned analysis was aimed at defining the spindle displacement caused by thermal deformations. The
influence of the rotor deformation caused by an external load and water pressure in bearing recesses did not

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consider this stage. The temperature field on the spindle surfaces obtained at the first stage was used as theboundary conditions of a finite element problem. The initial solid temperature (Ts=22.88 °C), defined at the firststage when flow simulation, is further used as the reference temperature at zero thermal strain for thermo-elastic
analysis. It is established that the spindle rotor ends elongate by around 2.7 µm due to the influence of thermal
deformations. Consequently, the pointed rotor deformation affects the positioning accuracy of a clamping device,
thereby deteriorating the machining accuracy of linear dimensions of a workpiece.

(a)(b)
Figure 2. Fluid Temperature Distribution on the Bearing Surfaces of the Casing (a) and Rotor (b)The thermal deformation of bearing surfaces results in the change of initial design gaps for the journal and thrustbearings. A thrust gap variation resulting from the thermal deformation is shown in Fig. 3, (a). The gap variationfrom the initial (design) gap size of 10 µm is represented in a circumferential direction along two diameters of thethrust bearing surfaces: the lowest (inner) and the highest (outer) positions. It is established that the highest thermal
displacements are taken place at the thrust inner diameter. On average, the mentioned gap size decreases by around
14 % compared to its design value. Minor deviations up to 0.12 µm of the thrust gap from average values (dashedlines in Fig. 3, (a)) are observed along the bearing circumferential direction at both inner and outer positions. On
the contrary, in the radial bearing direction, the average gap variation picks 0.26 µm as depicted in Fig.3, (a).
10

14along the bearing right egde

13

9.5

average gap variation in the radial direction
along the thrust outer diameter90.26 µm
8.5

along the thrust inner diameter
0

30

60

90

Gap size [µm]Gap size [µm]initial thrust gap (without deformation)

gap variation in
the axial direction

12

initial radial gap(without deformation)

11

average gap
variation

along the bearing left egde10

120 150 180 210 240 270 300 330 360

Bearing circumferential direction [Deg]

0

30

60

90

120 150 180 210 240 270 300 330 360

Bearing circumferential direction [Deg](a)(b)
Figure 3. Gap Variation in the Thrust (a) and Journal (b) Bearings Caused by Thermal Deformations

The thermal deformation of bearing surfaces also leads to the significant variation of the radial gap in the journalbearing (Fig. 3, (b)). Unlike the deformation pattern in the thrust bearing, the conjugated journal surfaces' thermal
displacements cause an alternate decrease and increase in the radial gap against the initial (design) gap value of
12 µm. Note that the average gap variation in the axial direction in the journal bearing is relatively minor, reachinga value of 0.22 µm. The primary influence on the bearing gap size for both the thrust and journal bearings iscaused by the thermal deformation of the bearing casing. Moreover, the deformation causes the form deviation of
bearing surfaces: flatness deviation for thrust bearing surfaces and roundness deviation for journal surfaces.

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INFLUENCE OF THERMAL DEFORMATION ON SPINDLE CHARACTERISTICS
This section presents the influence of bearing gap variation on a spindle film force and stiffness.
As shown in Fig. 4, (a), the thrust bearing surface's thermal deformation causes a decrease in both bearing film
force and stiffness. This phenomenon is due to the bearing surfaces' thermal expansion, causing a decrease in the
gap and, consequently, to an increase in recess pressure. While a supply pressure is keeping constant, this results
in increasing the pressure ratio β between supply 𝑝𝑝𝑠𝑠 and recess pressure 𝑝𝑝𝑟𝑟 ( β = 𝑝𝑝𝑟𝑟 ⁄𝑝𝑝𝑠𝑠 ) and thus causes a
decrease in the film force and stiffness for an opposed-pad thrust bearing.
3

1

1200

6000

4

800

2

4000

400
0

2000

0

1

2

3

4

5

6

7

8

9

Film stiffness [N/µm]

8000

Film force [N]Film stiffness [N/µm]

500

1600

0

5

450

6

400

350

300

0

2

4

6

8

Displacement of the thrust bearing [µm]Displacement of the journal bearing [µm]

(a)

(b)

10

Figure 4. Thrust Bearing Film Force (curves 1, 2) and Stiffness (curves 3, 4) vs Thrust Bearing Displacement (a) and
Journal Bearing Film Stiffness vs Journal Bearing Displacement (curves 5, 6) (b): Solid Curves Correspond to a Bearing
without Thermal Deformation; Dashed Curves – Bearing Surfaces Subjected to Thermal Deformation

The effect of non-uniformity of the gap on the journal bearing's output characteristics was then investigated. The
non-uniformity of the radial gap leads to the redistribution of pressures in bearing recesses and forming the change
in a resultant force supporting a rotor. As a result, the rotor is shifted to a new equilibrium position along thevertical axis by 1.726 µm under the influence of the unbalanced film force. Accordingly, this affects the accuracyof machining diametrical dimensions of the workpiece. Then, the influence of thermal deformation on the stiffnessof the journal bearing is investigated and presented in Fig. 4, (b). The thermal effect decreases the stiffness of thejournal bearing, but a comparative glance between curves 5 and 6 shows only a little quantitative difference in the
stiffness. For example, in the range of the journal bearing displacement up to 1 µm, a decrease in stiffness does
not exceed 0.5%.

CONCLUSION
The present study considered the influence of thermal effect on a machine tool spindle with water-lubricatedhydrostatic bearings. The study shows that when thermal effects account for, both the film force and stiffness
decrease. However, this trend is quantitatively different for spindle bearings: it is relatively minor for the recessedjournal bearings, and it causes the stiffness drop up to 14% for the opposed-pad thrust bearings. This also
contributes to the decrease in the accuracy of linear dimensions of machined parts.

ACKNOWLEDGMENTS
The presented results are a part of projects Invitational Fellowships for Research in Japan and Grant-in-Aid for
Scientific Research (C) of the Japan Society for the Promotion of Science.

REFERENCES
1. WARDLE F., Ultra Precision Bearings, 1st ed. Woodhead Publishing, 2015.
2. FEDORYNENKO D, et al., Increasing of Energy Efficiency of Spindles with Fluid Bearings, Acta Mechanica et
Automatica, 2017, Vol. 13, No. 3(41), pp. 204-209.
3. NAKAO Y., et al., Feasibility Study on Design of Spindle Supported by High-Stiffness Water Hydrostatic Thrust Bearing,International Journal of Automation Technology, 2014, Vol. 8(4), pp. 530-538.
4. FEDORYNENKO D, at al., Dynamic characteristics of spindle with water-lubricated hydrostatic bearings for ultraprecision machine tools, Precision Engineering, 2020, Vol. 63, pp. 187-196.
5. NAKAO Y, et al., Thermal Characteristics of Spindle Supported with Water-Lubricated Hydrostatic Bearings, Int. J.
Automation Technol., 2019, Vol.13, No.5, pp. 602-609.

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Multidisciplinary Design Optimization of a Tortuous Path Trim for a
Labyrinth Control Valve
Runlin GAN*, Xukang LI, Song LIU, Baoren LI*

FESTO pneumatic Technology Center, Huazhong University of Science & Technology,
Luoyu Road 1037, Wuhan, 430074, China(E-mail:grlhappy@126.com )Abstract. The labyrinth control valve has the advantages of low noise, low vibration, low flow velocity, goodcavitation resistance, and long maintenance intervals. It is widely used in petrochemical, nuclear energy and electric
power industries. A multi-disciplinary and multi-objective optimization design framework for the realization oflabyrinth flow control valve discs is proposed, and the results of flow field analysis and stress analysis are used as
design evaluation index. Based on the design premise of the flow rate, the optimization of the system aims to reducethe maximum velocity of the media and the stress of the disc. From the perspective of fluid and structure, the presented
method has important practical significance to improve the research level of labyrinth flow control valves.
Keywords: Multidisciplinary optimization, NSGA-II, Labyrinth valve

INTRODUCTION
In the industrial automation fluid control system, the control valve is one of the widely used fluid control
equipment that used to adjust the system parameters such as flow, pressure and temperature. The work
performance directly affects the production quality of the product and the benefit of the enterprise, and has greatsignificance in the entire industrial process system. With the technology and the large-scale industrial production
field improvement, enterprises have continuously expanded their demand for control valves. The industrial
process parameters have increased accordingly. The labyrinth control valve has put forward requirements for highpressure drop, high temperature, high pressure, low vibration and low noise [1-2].
Up to now, there are quite a few documents related to the improvement of the structure of the labyrinth flow
control valve. For example, Miller et al. [3] determined the outlet kinetic energy of the throttle plate as animportant parameter for evaluation the flow control valve based on engineer practice experience, and believedthat the outlet kinetic energy of the labyrinth channel should not exceed 275kpa in multi-phase flow. Tae-sig etal. [4] analyzed the flow characteristics in a single labyrinth channel at different expansion angles. The results
showed that the larger the rotation angle, the greater the pressure gradient. Na et al. [5] reduced the flow resistanceand increased the flow coefficient through added connect grooves in the labyrinth disc.
The relationship between the various parameters is coupled with each other. Univariate analysis takes a long timeand inefficient, and it is impossible to obtain the optimal structure and performance parameters from a global
perspective. Therefore, relevant scholars have conducted the research on multi-objective optimization. Corbera etal. [6] optimized the structure of the butterfly valve based on topology optimization and genetic algorithm toobtain a light butterfly valve disc that satisfies the structural safety and flow requirements. Laudon et al. [7]proposed a method to optimize the flow of the Tesla valve. The design was optimized through determined thecomplete optimization design parameters of the valve and combined the related optimization algorithms. Basedon the computer-aided and multi-objective optimization method, Ganesh et al. [8] carried out a parameterized
design on the valve plate of the axial piston pump, the fluid noise and structural noise was reduced, and theeffectiveness of the optimization process was proved through experiments. It is necessary to study the multiobjective optimization method of the labyrinth valve to obtain better flow field and mechanical performance.
This paper proposes the application of an optimization method, based on the NSGA-II algorithm for the multiobjective optimization solution of the flow velocity and stress of the labyrinth control valve under the restrictionsof the flow rate and the structure sizes. In order to achieve the goal, the main influence parameters that affect theflow characteristics and stress characteristics of the disc is analyzed, and parametric modeling is carried out. Then,
combined with CFD and FEM software to analyze the flow field and stress field respectively. Based on NSGA-II
algorithm to optimize the target.

MATHEMATICAL MODEL
To describe the flow phenomenon in the labyrinth valve, the standard k-ε turbulence model and the incompressible
flow model are adopted in ANSYS Fluent. The continuity, momentum equations to be solved are:Continuity equation:

1

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(  )+ ∇ • ( U) = 0.
t1)
In the formula, ρ is the density, kg/m3. U is the velocity vector. ∇ is the Hamilton operator.
Momentum equation:

( ui )p ( xi ) ( yi ) ( zi )+  • ( ui u ) = −++++ f i .
t
i
x
y
z

2)

In the formula, p represents the static pressure of the medium, Pa. fi represents the force of gravity along diverse
directions, m/s2. τ represents the viscous tensor, Pa.
The Zwart-Gerber-Belamri cavitation model based on transport-equation is used in numerical simulation. The
mass transport equation is given as follows: (v )+  • (v U ) = Re − Rc .
t3)

3 nuc (1 −  v )  v 2 Pv − P 21
Re = Fv
(
) .
RB
3 l

4)

3 v (1 −  )  v 2 Pv − P 21
Rc = Fc
(
) .
RB
3 l

5)

Where Re and Rc are the mass transfer source terms between the liquid and vapor phase in cavitation. αnuc is the⃗ is the velocity of the vapor phase.
nucleation site volume fraction. ρl and ρv are the density of liquid and vapor. 𝑈
RB is the radius of bubble. Pv represents the saturated liquid vapor pressure in 300K and P is the local fluid pressure.
Fv and Fc represent the evaporation and condensation coefficient in cavitation.

Figure 1. The structure of the labyrinth control valve

COMPUTATIONAL MODEL
Since the valve core of labyrinth control valve bears the main pressure drop and plays a major role in the fluidpressure and velocity variations, only the inlet and outlet accessories of the labyrinth disc flow channels are
considered in the calculation. Considering the symmetry of the valve plate, 1/8 of the flow channel in Fig.1 isselected as the calculation model.
The inlet surface and the outlet surface are provided as smooth and no-slip walls. The working conditions of fluid
analysis are shown in Table.1. Considering the phenomenon of cavitation, the medium is composed of water andwater vapor. The inlet pressure is 10.0MPa and outlet pressure is 2.0MPa. The temperature is 300K and there is
no heat transfer. Standard k-ε turbulence model is adopted. The material of the labyrinth disc is stainless steel and2

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the yield stress is 250MPa. Reference to the industrial valve pressure test (ISO5028), apply a pressure of 15MPa
to all positions on the inner wall of the valve, and limit the axial displacement of the labyrinth disc.
The grids of these areas have been divided, as shown in Fig.2. The number of grids affects calculation accuracy
and efficiency, and the grid sensibility analysis is performed. When mesh number varies from 101,232 to
1,223,599, the flow rate of the labyrinth control valve is calculated. It can be seen from Fig.3 that the mesh number
766,722 is appropriate.
Table 1. The working conditions
Working condition
Valve diameter (mm)100
Upstream pressure (MPa)10.0
Downstream pressure (MPa)2.0
Temperature (K)300
Flow rate (m3/h)≥150Phase 1 medium
Water
Phase 1 density (kg/m3)
998.2
Phase 1 dynamic viscosity (kg/(m·s))
0.001003
Phase 2 medium
Water-vapor
Phase 2 density (kg/m3)
0.5542Phase 2 dynamic viscosity (kg/(m·s))
0.0000134
Stainless steel
Young modulus
189GPaYield stress
270MPaFigure 2. Mesh of the labyrinth control valve

Figure 3. Grid sensibility analysis

DESIGN DESCRIPTION
Miller et al. [3] proposed the valve trim outlet speed and kinetic energy limit criteria, the multi-phase flow outlet
kinetic energy under various working conditions cannot exceed 275kpa. Otherwise noise and vibration will be
generated, so the fluid speed is selected as the design variable one. At the same time, considering the influence ofthe labyrinth channel structure on the safety of the disc structure design, the stress of the disc is taken as anotherdesign variable. In this paper, the maximum velocity of the flow channel and the stress of the structure are
optimized simultaneously under the constraints of flow rate and structure sizes. In this case, the NSGA-II
optimization algorithm is used for path optimization. The optimization problem of the labyrinth control valve canbe defined as the following equation:

 Minimized ：V(x),S(x)Subjected：Q( x)  Flowratex  C ( x)6)Where V(x) is the maximum velocity in the fluid path, S(x) is the maximum stress in the labyrinth channel structure,
and Q(x) is the flow rate. x =[x1, x2, ... , xn] is the variable element group, and C(x) is the value range.
For incompressible fluids, under the premise of constant flow rate, the flow velocity of the medium is inversely
proportional to the size of the flow area. The larger the flow area, the smaller the flow velocity. When the flow
area of the flow channel is constant, the effective flow area will be reduced and the flow velocity will increase3

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due to the "occupy" effect of the vortex in the flow channel. Therefore, the width of the labyrinth channel should
be expanded step by step along the flow direction, which can effectively control the speed of the incompressible
fluid.
The cross-sectional structure of a single labyrinth flow channel is shown in the Fig.4. The depth direction of theflow channel is perpendicular to the paper surface, and the depth is constant along the flow direction. The widthof the flow channel entrance is W1. The form expands step by step along the flow direction, and the width atnumber n is Wn=Wn-1×R. R is the expansion coefficient and n is the number of quarter turn. Meanwhile, H is thedepth of the flow. L1, L2, L3 are the length of the transition part, and L1=L2=L3. C1 is the size of the chamfer.
Considering the structure characteristics, the range of parameter values are defined in Table 2. In additional, thenumber of valve disc and the flow channel in each valve disc are also determined. There are 24 discs inside theinner core and each disc has 8 flow channels.

Figure 4. Chosen design parameters for flow path optimization
Table 2. Design parameters for flow path optimization
Parameters(mm)Initial value
Lower boundary Upper boundary Optimized value
1.264
1
2
1.40
W1
H
1.611
1
3
1.716
2.591
1
3
2.975
L1
0.255
0.1
0.5
0.305
C1
1.083
1
1.2
1.005
R

OPTIMIZATION ANALYSIS AND RESULTS
This article has two optimization goals, for multi-objective optimization problems, there is no unique optimal
solution in theory, but a set of optimal solutions called Pareto. According to the actual demand, select theappropriate result from the Pareto optimal solution.
The population size, number of iterations, number of variables, and the number of objective functions all affectthe optimization effect of the NSGA-II algorithm. The small number of populations and the iterations lead to local
optimal solutions and error solutions. A large number of populations and iterations improve the optimization effect.
However, it also leads to a decrease in calculation efficiency. In this paper, the population size is 30 and thenumber of iterations is 20. The first 40 sample points are obtained through the Optimal Space-Filling (OSF)
method. The maximum allowable Pareto percentage is 95 and the convergence stability percentage is 2.
Fig.5 is the evolution curve of the overall optimization iteration. It shows that after 10 iterations, the maximum
allowable Pareto percentage has reached 96.67, and the convergence stability percentage has reached 6.63.
Compared with the initial setting conditions, the number of iterations is saved by 50.70%.

Figure 5. The number of iterations (1. Maximum allowable Pareto percentage=95, 2. Pareto percentage, 3. Stability
percentage，4. Convergence stability percentage=2)

Fig.6 is the set of all solution sets. Based on the correlation Pareto optimal ranking, two candidate points are
selected. Candidate points 148, 278. The flow rate difference of the two candidate points is not big. The maximum

4

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velocity of candidate points 148 has a greater advantage, although its maximum stress value is slightly increased
compared to the candidate point 278. But at the temperature of 300K, the maximum yield strength of the disc is270MPa, and the safety factor of the candidate point is 3.26, which is much higher than the 1.5 value of the generaldesign. Within the stress range of the disc, the maximum velocity of the flow channel should be reduced as muchas possible to reduce vibration and noise. Therefore, candidate point 148 is selected as the optimization result
from the two candidate points.

Figure 6. Best two solutions of the optimization process and their position along the all results.

Figure 7. Comparison of the velocity between different valve plate design

Figure 8. Comparison of the stress between different valve plate design

Table 2 compares the original values with the optimized values. After the flow path optimization process, Fig.7
and 8 show the comparison between the initial design structure and the optimized design structure in terms of
fluid performance and disc stress distribution. Before optimization, the high flow velocity mainly appeared in the5

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first few turns of the labyrinth channel. At the same time, due to the large expansion coefficient R, it can be seenfrom Fig.7(a) that a considerable part of the flow channel has a low velocity, indicating that the flow in the flow
channel is uneven and the space utilization rate is insufficient. It can be seen from Fig.7(b), after optimized, theproportion of high velocity areas is significantly reduced. The flow in the back section of the runner is more
uniform. It can be seen from Fig.8(a) and (b) that the maximum stress basically appears at the corner of thelabyrinth channel, where stress concentration is easy to occur, and it needs to pay attention in design. Table 3
shows the performance improvement between the initial structure and the optimized structure. The flow rate is150.68m3/h, which meets the design flow rate. Compared with the initial structure flow rate 154.83m3/h, there islittle change. The maximum velocity is decreased from 104.10m/s to 96.54m/s, which is reduced by 7.26%. The
maximum stress is 82.20MPa. Compared with the initial design, the maximum stress is 89.71MPa, which isincreased by 8.37% after optimization.
Parameter
Flow rate (m3/h)
Velocity (m/s)
Stress (MPa)

Table 3. Objective improvement results
Initial design
Optimized design
154.83150.68104.10
96.54
89.71
82.20

Improvement (%)
-2.68
7.26
8.37

CONCLUSIONSIn order to obtain the best flow field parameters, a parametric model of the labyrinth valve is established. The
labyrinth valve system is decomposed into two goals of flow field and structural strength, which realizes theintegration of parametric modeling, finite element analysis and multi-objective optimization. The optimization
results show that, on the premise of ensure the flow rate, the maximum velocity of fluid path is decreased. At thesame time, the stress of the disc structure is reduced, and its safety and reliability are improved. From theperspective of fluid and structure, the realization of multidisciplinary design optimization and multi-objective
have important practical significance to improve the research level of labyrinth flow control valves.

ACKNOWLEDGMENTS
The authors would like to acknowledge the support of FESTO Pneumatics Center in HUST.

REFERENCES
1. KWON W C., KIM G R., PARK S C., et al., Design of a tortuous path trim for a high-pressure turbine bypass valve,
Proceedings of the Institution of Mechanical Engineers Part E Journal of Process Mechanical Engineering, Vol.224, No.2,
2010, p.149-153.
2. YONEZAWA K., OGAWA R., OGI K., et al., Flow-induced vibration of a steam control valve, Journal of Fluids &Structures, Vol.35, 2012, p.76-88.
3. MILLER HL., STRATTON LR. Fluid Kinetic Energy as a Selection Criteria for Control Valves, ASME Fluids
Engineering Division Summer Meeting, 1997, pp.1-8.
4. TAE-SIG G., KUISOON K. The Effect of Divergence Angle on the Control Valve Trim Characteristics, The KSFM
Journal of Fluid Machinery, Vol.16, No.1, 2013, p.32-39.
5. LV Z., Na Z., JING H., et al., Study on the improvement of flow characteristics for the optimal design of minimum flow
valve with labyrinth channel, 6th International Symposium on Fluid Machinery and Fluid Engineering, 2014, pp.1-4.
6. CORBERA S., OLAZAGOITIA J L., Lozano J A. Optimization of a butterfly valve disc using 3D topology and genetic
algorithms. Structural and Multidisciplinary optimization, Vol.56, 2017, p.941-957.
7. TRUONG TQ., NGUYEN NT. Simulation and optimization of tesla valves. Nanotechnology Conference and Trade Show,
Vol.1, 2003, p.178-181.
8. SEENIRAJ G K., IVANTYSYNOVA M. A Multi-Parameter Multi-Objective Approach to Reduce Pump Noise
Generation, International Journal of Fluid Power, Vol.12, No.1, 2011, p.7-17.

6

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Design of Magnetostrictive Power Generation Device from Pulsating
Pressure in Hydraulic Pipeline by Using Water Hydraulic Cylinder
Kaito MIYASHITA*, Shouichiro IIO**,
Tsuyoki TAYAMA***, Ryuichi ONODERA***, Shota ABE****

Graduate School of Science and Engineering, Shinshu University
Department of Mechanical Systems Engineering, Shinshu University
4-17-11 Wakasato, Nagano, 380-8553 Japan(E-mail: 20w4065c@shinshu-u.ac.jp)***
Tohoku Steel Co., Ltd.
23 Nishigaoka, Murata, Shibata, Miyagi, 989-1393 Japan**

Abstract. This study focuses on a magnetostrictive power generator that converts pulsating pressure in a water hydrauliccircuit to electric power through mechanical vibrations to suppress pressure pulsation. This study uses a water hydrauliccylinder as a pressure to force converter. Fluctuation force on the cylinder rod acts on a magnetostrictive piece. The authors
measured the pressure, flow rate passing water, and the thrust force on a rod. The pressure fluctuation frequency and theinitial cylinder volume is the significant parameter. As a result, the pressure absorption characteristics strongly depend on thepressure pulsation frequency and the initial volume. Several magnetostrictive generators designed and investigated themechanical strength by structural analysis.
Keywords: Pressure Pulsation, Water hydraulics, Vibration, Generator, Performance

INTRODUCTION
Pressure pulsation often occurs in fluid power systems, causing vibration, noise, and reduced life of connected
equipment. Accumulators are commonly used as pressure pulsation suppression devices in water hydrauliccircuits, and several other pulsation reduction devices had been proposed1), 2). All of these devices suppress
pulsation by dissipating the energy of pressure pulsation into the atmosphere as thermal energy or elastic strainenergy. Suppose both suppressions of pressure pulsation and electric power generation can achieve. In that case,it will be possible to supply power to sensors and other devices used in pipelines, and it expects to apply to the
IoT, which has been attracting attention in recent years. Therefore, the authors have proposed a Pulsation
Absorbing Device (PAD) that absorbs pressure pulsation in a water hydraulic circuit and generates electricity,and have investigated its basic characteristics3), 4). Since the energy loss of the conventional diaphragm type
PAD is significant and the transmission of vibration to the power generation device is not easy, the authors start
to investigate the magnetostrictive power generation device using a water hydraulic cylinder. In this paper, theauthors report the results of the preliminary performance tests for the realization of a power generation device
using a water hydraulic cylinder type PAD. In particular, the frequency of the pressure pulsation in the circuitand the effect of the internal volume of the water hydraulic cylinder type PAD on the pressure pulsation
absorption characteristics report. In addition, the characteristics of the force generated in the cylinder rod
investigate. Besides’ the authors reports investigating results of the optimal shape of magnetostrictive materials
for magnetostrictive power generation based on structural analysis.

EXPRTIMENTS AND ANALYSIS
Experimental Equipment and Methods
The authors investigate the effects of the frequency of the pressure pulsation in the circuit and the initial volumeof the water hydraulic cylinder on the pressure pulsation absorption characteristics and the characteristics of theforce generated on the cylinder rod experimentally. Figure 1 shows the water hydraulic circuit. Table 1(a) showsthe experimental conditions when the initial volume is constant and the pump rotation speed change to confirm
the pulsation absorption performance when the pulsation frequency change. Table 1(b) shows the experimental
conditions when the pump rotation speed is constant and the initial volume change to confirm the pulsation
absorption performance when the initial volume change. The working fluid is tap water. Swash-plate type piston
pump unit (KYB Corp., 16MPa max., 22.5L/min@1500rpm), six cylinders, use. The water from the pumpregulate to the desired pressure by a relief valve and passed through an upstream High response transducer(JTEKT, PMS-8M-2), a water hydraulic cylinder type PAD, a downstream High response transducers (JTEKT,PMS-8M-2), an ultrasonic flow meter, a needle valve, a chiller unit(ORION, RKS1503J-MV-01000), and afilter before returning to the tank. Semiconductor pressure sensors (KEYENCE, upstream side: GP-M100,

1

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downstream side: GP-M10) measure the other two pressures between a needle valve . For the piping of theexperimental circuit, 10A Sch/80 stainless steel pipes use and connect in a straight line at the length shown in
the figure. Connect the water Hydraulic cylinder type PAD to the pipe with a T-junction so that Fill the cylinder
piston side with working water. The cylinder stroke can fix at any position by installing a fixed wall at the rod
end. The initial volume defines as the volume on the piston side of the cylinder. When a pressure pulsation actson the piston, a thrust force generate in the rod. In magnetostrictive power generation, a magnetic field generate
when the magnetostrictive material distort, which is then used to generate electricity. So that the force sensor(small compression type load cell, Kyowa Dengyo, LMC-A-20KN) by install between the rod tip and the fixe
wall measure the force acting on the magnetostrictive material. The rod side of the cylinder is opened to theatmosphere. Using the data logger record the pressure, flow rate, and force values a sampling frequency of 1
kHz for 5 seconds.

Fose senser

Table 1 Experimental condition

Figure 1. Testing water hydraulic circuit.
Table 1. Experimental condition.
.
(a) When the initial volume is constant Initial volume.
(b) When the pump rotation speed is constant.

Fluid medium

Tap water

Fluid medium

Tap water

Fluid temperature[℃]̅̅̅̅Average pressure𝑃
up [MPa]̅̅̅
Average pressure 𝑃2 [MPa]20

Fluid temperature [℃]
Average pressure ̅̅̅̅𝑃up [MPa]

20

Pump rotation speed[rpm]
PAD
Initial volume V0×10-3 [m3]

3.2
0.715
1200, 1400, 1600,
1800
0.467

Average pressure ̅̅̅
𝑃2 [MPa]
Pump rotation speed[rpm]
PAD
Initial volume V0×10-3 [m3]3.2
Varies with pumprotation speed
1800
0.62, 0.467, 0.317,
0.166, 0.0015

Analysis Conditions and Methods
First, Figure 2 shows how the model introduces into the actual machine. In the future, the authors will refer toFigure 2(a) with the neck as Model A and Figure 2(b) with the notch as Model B. In addition, Figure 2(c) shows
the state of the manufactured power generation device.In this figure, the magnetostrictive material shows in red,the coil wind as shown in black, and the magnet attaches as shown in blue. In this device, the magnetostrictive
material strains to generate a magnetic field, and the magnets and coils is attached to the device induce
electromagnetic induction to generate electricity. In the actual machine the magnetostrictive material will notfixe but only sandwiched between the cylinder cover and cylinder rod.
The optimal shape of the magnetostrictive material must incorporate elements that make it easy to wind coils
and easy to attach magnets. Hence, the author conducts a structural analysis to find a new shape. Table 2 shows
the physical properties of the magnetostrictive material. As the analysis software use the inventor, and the meshsize set to 0.01 of the model size ratio for both analysis time and accuracy. The load set to 17,000 N, which isslightly higher than the maximum load, in order to find a shape that would not fracture even if the maximum
load, which has been confirmed in a water hydraulic cylinder type PAD, was applied. The restraint conditions
fixe restraint, and magnetostrictive materials use.

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Magnetostrictive
material
φ10×30[mm]CoilWire diameter
φ0.05[mm]Resistance12.3[kΩ]

Magnet0.5×0.5×1[mm](a) model A.

(b) model B.

(c) Photo of the testing power generation device.
Figure 2. Materials used in the experiment.
Table 2. physical properties of magnetostrictive material.
205[GPa]
0.33
77[GPa]397[MPa]
659[MPa]
8.41[g/cm3]

Young’s modulus
Poisson’s ratio
Shearing modulus
0.2% proof stressTensile strength
DensityDISCUSSION OF THE EXPERIMENTAL RESULTS AND ANALYTICAL RESULTS
Thrust of the Cylinder Rod
Figure 3 shows the experimental results when the initial volume is kept constant. This figure shows versus pumprotation speed for the pressure pulsation absorption efficiency η and the difference between the p-p values ofthrust force ΔF. η is the difference between the upstream energy and the downstream energy divided by the
upstream energy. Figure 3(a) shows that η increases with the pump rotation speed up to 1600 rpm, and at 1800
rpm, the value is almost the same as at 1600 rpm. Figure 3(b) shows that ΔF reaches its maximum value at 1600rpm and decreases after that, but there is no significant difference in ΔF between 1600 rpm and 1800 rpm. These
results suggest that in the pump rotation speed range in which this experiment was conducted, the efficiency ishigh at high pump rotation speeds of 1600 rpm to 1800 rpm. The force acting on the magnetostrictive material
may be significant.

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Figure 4 shows the experimental results when the pump rotation speed keep constant. This figure shows versus
initial volume for the pressure pulsation absorption efficiency η and the difference between the p-p values ofthrust force ΔF. From Figure 4(a), the authors can see that η increases with V0. On the other hand, figure 4(b)shows that ΔF decreases as V0 increases. These results show that there is a trade-off between η and ΔF. Thisphenomenon occurs because the water hydraulic cylinder type PAD does not absorb the pressure pulsation onlyby the vibration of the rod, but also by the vibration of the water, molecules in the cylinder are elastic andabsorb the pressure pulsation at the same time as the rod. Therefore, when the water in the cylinder increased
due to the initial volume increase, the ratio of water in the cylinder absorbing pressure pulsation increased, andthe ratio of pressure pulsation transmitted to the rod decreased, so ΔF decreased and η increased.

(a) Pressure pulsation absorbing efficiency.
(b) Differential thrust force ΔF.
Figure 3. pulsation absorbing performance versus Pump rotation speed. (V0=0.467×10-3[m2])(a) Pressure pulsation absorbing efficiency.
(b) Peak to peak of thrust force ΔF.
Figure 4. pulsation absorbing performance versus initial volume. (1800[rpm])

Investigating of the Magnetostrictive Material ShapeIn the present analysis, the authors investigated the neck model, which is easier to strain and winding a coil, andthe notch model, which is easier to attach magnets, and obtained two types of shapes. First, figure 5 shows theresults of the analysis for the mesh and von Mises stress of Model A. In order to increase the strength of themodel, there are one notch, one in the load part and one in the fixed part, and these two notches are located at
180 degrees symmetry. From this figure, the stress considers concentrating at the neck, and the maximum vonMises is about 350 MPa. The authors believe that this value is sufficient to strain the magnetostrictive material,and from the standpoint of safety factor, no rupture will occur, so practical application is possible.
Next, the authors conduct a structural analysis using a simple structure with only a notch, in order toexperimentally confirm the properties of the magnetostrictive material in a simple shape. Figure 6 shows theresults of the analysis for the mesh and von Mises stress of Model B. The maximum von Mises stress is about
150 MPa, which is also sufficient to strain the magnetostrictive material, and from the viewpoint of safety
factor, it is possible to put it into practical use because it is thought that rupture will not occur.

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(b) von Mises stress.
Figure 5. Analysis results of model A.

(a) Mesh.

(b) von Mises stress.
Figure 6. Analysis result of model B.

CONCLUSION
The authors experimentally investigate the relationship between pressure pulsation absorption efficiency and rod
thrust force in a new type of PAD using a water hydraulic cylinder. As a result, the authors confirm theabsorption effect of pressure pulsation, and show the relationship between pressure fluctuation waveform andthrust force. These characteristics find to be strongly dependent on the initial volume and pump rotation speed.
Also’ based on the relationship between the pressure pulsation absorption efficiency and the thrust force,
structural analysis of the magnetostrictive material was performed to realize the power generation device, theoptimum shape of the magnetostrictive material was determined. The results of the power generation test willreport at the meeting.

REFERENCES
1.
2.
3.
4.

Ken Ichiryu, “Development of accumulator (Hi-Pulsober) for high-frequency ripple absorption of hydraulic system,”
Bulletin of JSME, Vol. 15, No. 88, pp.1215-1227, 1972.
Yasuo Sakurai, Norikazu Hyodo, and Kenichi Aiba, “Component to reduce pressure pulsation,” Japan Fluid Power
System Society, pp. 41-44, April 2017.
K. Mori, S. Iio：Feasibility Study on a Pulsation Absorber for Vibration Power Generation, Proc. the 8th International
Conference on Fluid Power and Mechatronics, FPM2019-408(2019), pp.1-3.
K. Mori, S. Iio ： Development of a Liquid Flow Pulsation Damper which Works as Power Generator, The
Proceedings on Spring Conference of Japan Fluid Power System Society 2019 （2019）, pp.95-97.

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Flow Characteristics of a Cavitating Jet through a Small RectangularOrifice with Different Aspect Ratios
Hironori TAKEI*, Kohei TERAKAWA*, Shouichiro IIO**,
Kotaro TAKAMURE***, Tomomi UCHIYAMA***, Futoshi YOSHIDA*****

Graduate School of Science and Technology, Shinshu University
Department of Mechanical Systems Engineering, Shinshu University
4-17-1 Wakasato, Nagano-shi, Nagano, 380-8553 Japan(E-mail: shouiio@shinshu-u.ac.jp)***
Institute of Materials and Systems for Sustainability, Nagoya University
Hurou-cho, Tikusa-ku, Nagoya-shi, Aiti, 464-8601 Japan****KYB Corporation
1-12-1, Asamizodai, Minami-ku, Sagamihara-shi, Kanagawa, 252-032 Japan**

Abstract. Cavitation is one of the significant issues for water hydraulic components. The authors focus on a cavitating
jet flow and jet noise characteristics from a rectangular orifice for various aspect ratios. This study deals with thecavitation region in the jet to reveal the relation between the cavitation and vortex structure issuing from a rectangularorifice. The experimental results show that the vortex in the jet shear layer corresponds with the cavitation generation
region. No cavitation occurs in the jet center region just downstream of the exit. The cavitation bubbles grow anddissipated at almost the same normalized downstream position for the lower aspect ratio orifices. The cavitating planar
jet behavior differs from those for the other lower aspect ratios. The lower aspect ratio jet shows axis switching.
Keywords: Cavitation, Jet, Rectangular Orifice, Water Hydraulics, Visualization

NOMENCLATURE
AR
C
De
DO

: jet exit aspect ratio [-]: flow coefficient [-]: equivalent diameter [mm]: disolved oxigen concentration [ppm]

P1
P2
Re
x
σ

: upstream pressure [×106 Pa]: downstream pressure [×106 Pa]: Reynolds number [-]: distance from orifice [mm]: Cavitation number [-]INTRODUCTION
Aqua drive system (ADS), which uses tap water as the working fluid, has recently been attracting attention as aclean and environmentally friendly system (1)-(3). ADS is superior to hydraulic systems in terms of hygiene andenvironment, compliant with various standards such as HACCP and ISO 22000. For this reason, ADS isincreasingly being used in manufacturing processes that require cleanliness, such as food, pharmaceuticals, andsemiconductors, and in equipment that requires environmental compatibility, such as elevators for nursing careand gates for storm surge protection. However, due to the physical properties of water, the working fluid in ADS
is more prone to cavitation than hydraulic oil. So, noise and vibration problems occur apparently.
The previous researches by the authors showed that the cavitation phenomena in a water hydraulic spool valve
depend on the width and the opening area of the control orifice and the differential pressure between the orifice.
Furthermore, the variation of vibration and noise with downstream pressure shows a peak when the pressure
becomes around 0.1 MPa(4). The control orifice in a water hydraulic spool valve is rectangular, and the width andaspect ratio of the orifice changes simultaneously. Therefore, it is essential to check the dependency of cavitationon the width or the aspect ratio individually. The authors made and used a small rectangular orifice, the area isapproximately 1.6mm, for evaluating these influences independently. Although some previous studies onrectangular orifices, such as micro-orifices(5), there is insufficient knowledge on the cavitating jet, whose
conditions of actual orifice size, Reynolds number, and cavitation number are in this study. The orifice area isconstant and almost the same as the control port area of the spool valve. Thus, the experiment parameters are
aspect ratio, differential pressure, downstream pressure, and dissolved oxygen concentration. The authors reported
that the orifice aspect ratio significantly influences the origin of the cavitation jet, and for a higher aspect ratio,1

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The 11th JFPS International Symposium on Fluid Power 2020the less cavitation and noise(6). However, the detail of the cavitation jet is unknown. Therefore, this study aims tograsp the origin and growth of the cavitation jet by visualization.

EXPERIMENTAL EQUIPMENT AND METHOD
Figure 1 shows the testing water hydraulic circuit and the arrangement of visualization devices. The working
water pressures, P1, P2, are monitored by pressure transducers and adjusted by relief valves. The water passes anaccumulator for pulsation suppression and into the test section and returns to a reserver. The water temperature
was maintained constant by a chiller. Through the experiment, the dissolved oxygen concentration (DO) and thetemperature were measured at the inlet of the test section. The DO and temperature range from 6.43 to 7.06 ppm
and 22.3 to 22.5 °C, respectively. Table 1 lists experimental conditions.
Figure 2 illustrates a schematic view of the test section. The section is a transparent acrylic vessel to observe thecavitation jet directly. Stainless steel plates sandwich the vessel for easy connection to the pipe. An orifice plateis screwed at the center of the bottom plate, as shown in Fig. 2. Figure 3 shows four types of orifice plates usedin this study. A type has a circular hole of 1.46 mm diameter, and the other types with rectangular openings withaspect ratios of AR=1.00, 6.76, and 52.0. All plates have the same opening area of 1.69mm2.
The backlighting method with a halogen lamp captures the whole jet image viewing from the front of the test
section. On the other hand, the laser light sheet with CW laser (Kanomax, CW532-10-3W) can slice the cavitating
jet to grasp cross-sectional images viewing from the top of the test section. A high-speed camera is FASTCAM
SA-X2, Photron with a lens (Nikon, NIKKOR 50mm f/1.4). The frame rate of the high-speed camera was 30 kfps,
the power of the halogen lamp was 500 W. The time-averaged images were using Image-J software.

Figure 1. Experimental Circuit

Figure 2. Test Section

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1 .

9.36 mm×0.18 mm(d) AR=52.0.

1.30 mm×1.30 mm
3.38 mm×0.50 mm(b) AR=1.00(c) AR=6.76
Figure 3. Orifice plates(a) φ1.46 mm

4(unit in mm.)Figure 4. Cross-section of the orifice plate
Table 1. Experimental condition

Fluid medium
Temperature [deg. C.]
Upstream pressure P1 [×106 Pa]
Downstream pressure P2 [×106 Pa]
Disolved Oxigen concentration DO [ppm]
Orifice types

Tap water
22.3～22.5
0.50～3.00
0.00
6.43～7.06
φ1.46mm (circle)
AR=1.00, 6.76, 52.0 (rectangle)

thickness is 0.5 mm

EXPERIMENTAL RESULTS AND DISCUSSION
Figure 5 shows the front views of the cavitating jet at different upstream pressures. We can easily recognize thatthe jet generates at the exit and dissipates downstream. The size of the jet is different between two pressure
conditions; cavitation jet forms near the jet exit for P1=0.5×106 Pa. In contrast, the jet grows until far from the exit
for P1=3.0×106 Pa. Compare the length of the jets between the exit condition. The length for AR=52.0 is shorter
than that for other exit shapes. The previous research focused on cavitation type based on bubble diameter. There
are two types of cavitation; large diameter bubble cavitation originates from the gas cavitation by the concentrationof dissolved gas in the water, and tiny cloud-like bubble by the local pressure drop(4)(6). The bubble cavitation
occurs at P1=3.00×106 Pa, especially in cases of φ1.46 mm, AR=1.00, and 6.76. In the case of AR=52.0 shown in
Fig. (d), the cavitation region is small around the jet exit.
Figure 6 illustrates instantaneous cross-sectional slicing images of the cavitating jet at the different downstream
positions. The brighter region is a higher density of cavitation bubbles. For the circular jet in Fig. (a), a ring shapecavitation can be observed at x/De≈0, and the ring is in the jet shear layer. This fact indicates the cavitation
produced by the intense shear stress. The circle shape bubble distribution is not uniform and fluctuating, that
confirmed from the movie. The ring shape becomes unclear at x/De≥3.43. The cavitation ring disappears far
downstream for the circular jet in Fig. (a). For the rectangular jet of AR=1.00 in Fig. (b), a diamond shapecavitation appears at x/De≈0. The shape rotates at approximately 45 degrees from the jet exit shape. Thus, rapid
axis switching will occur because of a rectangular orifice with a low aspect ratio(7)(9). the cavitation shape at
x/De≈0 for rectangular jets of AR=6.76, 52.0 shows the exact shape of each exit. The difference between thesecavitation patterns is that the jet of AR=52.0 keeps its shape far downstream, as shown in Fig. (d).
Time-averaged images of cross-sectional cavitation patterns are in Fig. 7. These images could get from an averageof a thousand images for 33msec. Thus, we can precisely catch the significant phenomenon mentioned above
shown in Fig. 6. For the circular jet in Fig. (a), the circle shape remains for x/De≤4.11, and cavitation bubbles do
not penetrate near the jet centerline, the x-axis. The cavitation region in the rectangular jet of AR=1.00 deforms,
and switching phenomena at x/De≈0 and covers on the jet center area at x/De=3.85, due to enhancement of jet
mixing(7)-(9). The cavitation distribution pattern in the jet of AR=6.76 in Fig.(c) appears typical deformation
process of a coherent structure of a non-cavitation jet. It means the deformation process due to the self-induced
velocity of a vortex ring. The corners of the vortex ring with a small curvature move faster downstream. These3

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cavitation patterns at x/De≈0 and 3.44 shown in Fig. (c) are similar with deformed vortex rings in a non-cavitation
jet from the same exit shape. The cavitation region in a two-dimensional jet of AR=52.0 does not penetrate the jet
center area regardless of the tiny gap.
Figure 8 illustrates the variations of flow coefficient with the change of cavitation number. The upstream pressureof P1=3.00×106 Pa was constant, and only the downstream pressure was changed. From the results, the flow
characteristics improve in the order of φ1.46mm, AR=1.00, AR=6.76, AR=52.0, almost independent on thecavitation number. As the cavitation number decreases, the flow coefficient starts decreasing at some cavitation
number drawn in the red circles in Fig. 8. The choked flow by the cavitation will happen from the point. The
equivalent diameter of the orifice is De=0.87mm for AR=6.76, De=1.30mm for AR=1.00, and De=1.46mm for
φ1.46mm. Under the conditions of AR=6.76, AR=1.00, and φ1.46mm, the larger the equivalent diameter shows thesmaller inception cavitation number. On the other hand, no significant cavitation blockage was observed for
AR=52.0. The reason for this phenomenon is under investigation.

CONCLUSION
Visualization of cavitating jets with different aspect ratios was performed in this study. The obtained cavitation
images indicated that the cavitation is produced in the jet shear layers. The jet exit shape significantly influences
cavitation formation and pressure-flow coefficient characteristics. The growth and collapse positions normalized
by the equivalent diameter of the cavitation in rectangular jets of AR=1.00, 6.76, φ1.46mm are almost the same.
Future work is to grasp the relation between the cavitation phenomenon and flow discharge characteristics.

P1=0.50×106 Pa
P1=3.00×106 Pa(a) φ1.46 mm

P1=0.50×106 Pa
P1=3.00×106 Pa(b) AR=1.00

P1=0.50×106 Pa
P1=3.00×106 Pa(c) AR=6.76

P1=0.50×106 Pa

P1=3.00×106 Pa(d) AR=52.0

Figure 5. Instantaneous front views along with the z-axis of a cavitating jet

4

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x/De≒0x/De=3.43

x/De=4.11(a) φ1.46 mm, Re=29900

x/De≒0x/De=3.08

x/De=3.85(b) AR=1.00, Re=28700

x/De≒0x/De=3.44

x/De=4.59(c) AR=6.76, Re=20900

x/De≒0x/De=5.66

x/De=8.49(d) AR=52.0, Re=11500
Figure 6. Instantaneous cross-sectional slicing images of a cavitating jet at the different downstream position

5

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x/De≒0x/De=3.43

x/De=4.11(a) φ1.46 mm, Re=29900

x/De≒0x/De=3.08

x/De=3.85(b) AR=1.00, Re=28700

x/De≒0x/De=3.44

x/De=4.59(c) AR=6.76, Re=20900

x/De≒0x/De=5.66

x/De=8.49(d) AR=52.0, Re=11500
Figure 7. Time-averaged cross-sectional images of a cavitating jet viewing from the downstream

6

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1.4

1.4

1.
1.

1.

1.

.

.

.

.

1.4

1.4

..

.

.
.

.
.

1

Figure 8. σ- C curves of orifices with various aspect ratios (P1=3.0×106 Pa)

ACKNOWLEDGMENT
This work was carried out by the joint research program of the Institute of Materials and Systems for Sustainability,
Nagoya University.

REFERENCES
1. Z. Y. Li, Z. Y. Yu, X. F. He, S. D. Yang, "The development and perspective of water hydraulics", Proc. JFPS International
Symposium on Fluid Power, 1999(4), 1999, pp.335–342.
2. E. Trostmann, Water Hydraulics Control Technology, CRC Press, 1995.
3. Progress of Aqua Drive technology (Japan Fluid Power System Society establishment 40th anniversary publication), The
Japan Fluid Power System Society, 2011.
4. H. Okabe, Y. Tanaka, A. Watanabe, F. Yoshida, S. Iio, Y. Haneda, "Cavitation in a spool valve for water hydraulics", IOP
Conf. Series: Earth and Environmental Science, 240, 062029, 2019.
5. S. Cassineri, et al., "Experiments on liquid flow through non-circular micro-orifices", Micromachines 2020, 11(5), 510,
2020.
6. A. Watanabe, H. Okabe, F. Yoshida, S. Iio, "On cavitating jet issuing from a rectangular orifice", Proc. the 8th International
Conference on Fluid Power and Mechatronics (FPM2019), Wuhan, FPM2019-406, 2019.
7. K. Toyoda, "Vortices in jets", Journal of Japan Society of Fluid Mechanics, 24, 2005, pp.151-160.
8. A. K. M. F. Hussain, "Coherent structures and turbulence", J. Fluid Mech., 173, 1986, pp.303-356.
9. K. Toyoda, A. K. M. F. Hussain, "A study of the vortical structures of noncircular jets", Transactions of the JSME Ser.B,
55(514), 1989, pp.1542-1545.

7

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Organized Session | Organized session

Aqua Drive 2[OS4-2-01] Switching Control of Latex Balloon Expansion by using Fluidic Switching
Valve mediated with Coanda Effect〇Keita Kaneko1, Kenjiro Takemura1 （1.Keio University）[OS4-2-02] Comparison of Model-Free Adaptive Displacement Control and Model
Predictive Displacement Control for Tap-Water-Driven Muscle
Considering Load Variation during Experiments〇Satoshi Tsuruhara1, Ryo Inada, Kazuhisa Ito1 （1.Shibaura Institute of Technology）[OS4-2-03] Experimental study on Dual-Layer Type Vortex Cup Driven by Aqua
Drive System〇Wataru Kobayashi1, Tetsuya Akagi1, Shujiro Dohta1 （1.Okayama University of Science）©The Japan Fluid Power System SocietyOS4-2-01

The 11th JFPS International Symposium on Fluid Power 2020

Switching Control of Latex Balloon Expansion
by using Fluidic Switching Valve mediated with Coanda EffectKeita KANEKO* and Kenjiro TAKEMURA***

Graduate School of Science and Technology, Keio University
3-14-1, Hiyoshi Kohoku-ku, Yokohama, Japan(E-mail: keitakanekomsazb@keio.jp)**
Department of Mechanical Engineering, Keio University
3-14-1, Hiyoshi Kohoku-ku, Yokohama, Japan(E-mail: takemura@mech.keio.ac.jp)

Abstract. Soft robots have advantages in terms of safety, softness and compliance. However, fluid driven soft actuators
often employed in soft robots require corresponding numbers of pressure supply/valve to drive. Here, we take noticeof a fluidic valve which can control flow without mechanically moving parts from the viewpoint of simplifying thedriving system of soft actuators. We develop a system which consists of a pump, a fluidic switching valve, and twolatex balloons to demonstrate the feasibility of introducing the fluid valve to soft robotics. Since the valve can switch
the flow between two outlets when the pressure difference between outlets is under several kPa, we have employed
the latex balloon connected to each outlet, which has a peak in its PV diagram. The system can control each balloon’s
expansion by switching the flow from the pump. Experimental results proved that the system could actuate eachballoon at will.
Keywords: Soft robot, Fluidics, Coanda effect.

INTRODUCTION
In recent years, there has been a demand for robots that can operate in the environment close to human. They
could be used as nursing care robots, entertainment robots, and customer service robots etc. [1]. Such robots are
expected to be safe enough against human to reduce possible physical harms when they become into contact withhumans, flexible enough to behave like living organisms, and compliant enough to cope with unexpected
situations and environments [2-4]. For these reason, soft robots, in which the structure and actuators themselves
are flexible, have been attracting wide attentions [5]. Soft robots are often constructed and actuated by means of
fluid-driven flexible rubber actuators [6]. Then, there arises a critical issue. When developing a multifunctional
soft robot, the mass and volume of the entire system increase because multiple pressure sources/valves are required
according to the number of degrees of freedom of motion required in the robot. A number of researches have beenreported on the application of fluidics technology to pneumatic soft robots, where multiple motions can begenerated from a single input, suggesting the possibility of fluidics to increase the number degree-of-freedom withsimple configuration [7-9]. Compared to pneumatically driven soft robots, hydraulic systems for soft robotics
have not been greatly explored yet. Many of the previously reported hydraulic soft robots employ rigid and bulky
fluid control components [10, 11].
In light of the above background, we focused on a fluidic valve, which is one of the important elements of fluidics.
The idea it to realize a multi-degree-of-freedom motions by appropriately switching the pressure flow generated
by a pressure source by using a simple fluidic switching valve. To demonstrate this concept, this study aims torealize a fluid-driven system that can selectively expand balloons by using a simple fluidic switching valve.
Main flow

Control flow 2

Outlet 1

Balloon 1

Throttle valve

Pump

Fluidic switching valve

Control flow 1

Outlet 2

Balloon 2

Figure 1. A schematic of the proposed drive system.

1

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SELECTIVE DRIVE OF SOFT ACTUATORS USING A FLUIDIC VALVE
Figure 1 shows a proposed drive system capable of selectively expand each of two balloons at will using a simple
fluidic switching valve. The proposed drive system consists of a fluidic switching valve that switches the directionof a main flow discharged from a pump by using control flow. A balloon is attached to each of the two outlet
channels. The main flow flows into one outlet according to the control flow input, and a throttle valve locateddownstream of the balloon increases the gauge pressure of the balloon to inflate the corresponding balloon. By
switching the control flow, the other balloon can be selectively inflated. By doing so, selective drive of multiple
balloons with a relatively simple configuration can be achieved, contributing to the downsizing of entire system.

FLUIDIC SWITCHING VALVE
Figure 2 shows the schematics of the fluidic switching valve, which has been considered as an important fluidic
element in fluidics. Fluidics intends to construct a logic circuit with fluid flow like an electrical circuit. For such
purpose, the Coanda effect in which fluid flow along a channel wall [12], has been employed to develop a fluidic
logic element. We take advantage of this digital switching function in driving soft actuators. The fluidic switchingvalve can switch the outlet flow which is discharged from the pump (main flow) by inputting control flow fromthe side (control flow). In other words, the main flow can be directed into the outlet 1 by inputting the control
flow 1, and vice versa (see Figure 2).

BALOONAs shown in Figure 2, directing the main flow to the outlet 1 (2) creates a difference in the discharge flow rate
between the outlets. This flow difference causes the internal pressure increase of the balloon connected to theoutlet 1 (2) thanks to the throttle valve (cf. Figure 1). Therefore, the difference in deformation between the two
balloons can be achieved by the pressure difference, ∆P, which depends on the difference in flow rate between
the outlets.
Figure 3 shows the PV diagram of a balloon with (a) linear and (b) nonlinear relation between internal pressureand volume. In the case of a balloon with linear PV diagram (a), ∆! must be large to produce a sufficient
deformation or expansion. On the other hand, a balloon with a peak in PV diagram (b) deforms little up to a
threshold pressure and deforms significantly with the threshold pressure and above. In other words, the balloons
having peaks in their PV diagram do not require high pressure for obtaining large deformation. Even if ∆! issmall, it is possible to generate a sufficient difference in balloons deformation. Therefore, a latex balloon (polka
dot balloon, Tiger Rubber Co., Ltd., Japan) was employed in this study.
(a)Control flow 2(b)Outlet 1

Outlet 2
Control flow 1
Main flow
Figure 2. A schematic of a fluidic switching valve. The main flow is directed to (a) the outlet 1, and (b) the outlet 2.

(a)

(b)
Inner pressure

Inner pressure

Threshold

VolumeVolumeFigure 3. A PV diagrams of a balloon.
(a) Linear PV diagram, (b) Nonlinear PV diagram with a peak.

2

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FABRICATION AND EVALUATION OF COMPONENTS
Fabrication of Fluidic Switching Valve
We developed a fluidic switching valve made of acrylic resign as shown in Figure 4. The diameters of main flow
inlet, control flow inlet, and outlet are 13 mm, 3 mm, and 3 mm, respectively. The cross-section of an orifice part
where the main flow and the control flow collide is a rectangle with width of 2 mm and thickness of 3 mm. The
employed working fluid is water.

Evaluation
For the evaluation of each component, (i) switching characteristics of the fabricated fluidic switching valve and(ii) PV diagram of the latex balloon were experimentally obtained.
Figure 5 shows the experimental setup for switching characterization. A positive displacement pump (V-15, Iwaki
Corporation, Japan) was connected to the main flow inlet of fluidic switching valve, and the control flow inlet 2
was closed. A syringe (SS-50ESZ, Terumo Corporation, Japan) was connected to the control flow inlet 1 to direct
the main flow to the outlet channel 1. A flowmeter (SEN-HZ06K, Uxcell, China) was connected to each outlet tomeasure the flow rate along each outlet channel. The flow rate of the main flow was varied from 0 to 25.7 mL/s
by 20 equal intervals. Also, the cross-section ratio of the throttle valve (KT-6, Cole-Parmer LLC, USA) connectedto the outlets was varied from 59% to 100% (fully closed) by 5 intervals. Figure 6 shows the flow rate differenceof the outlet 1 against the outlet 2 ("!"# 1 - "!"# 2). From this result, the volumetric flow rate discharged from theoutlet 1 was about 0.25 mL/s higher than that from the outlet 2 in almost all conditions. Note that, there is aleakage from the outlet 2.
(b)

ϕ5

(a)

2

ϕ13

ϕ15

6

ϕ5

ϕ3

ϕ3

t=3

(mm)

Figure 4. The developed fluidic switching valve.
(a) Cross-section and dimensions, (b) Actual view of the valve.

Pump
20 steps from 0 to 26.7 mL/s

Control flow 2
Close

Control flow 1
7.8 mL/s

Flowmeter 1 !$%& )Throttle valve
5 steps from 59 to 100%

Flowmeter 2 !$%& (Figure 5. A schematic of evaluation experiment of fluidic switching valve.

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flow rate difference (mL/s)

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1.5
Cross-section ratio 100%
Cross-section ratio 86%

1

Cross-section ratio 78%
0.5

Cross-section ratio 67%
Cross-section ratio 59%

0

Inner pressure (kPa)

0

5

10
15
20
25
Main flow's flow rate (mL/s)
Figure 6. The difference of flow rate of the outlet 1 against the outlet 2.

10
8
6
4
2
0
0

20
Volume (mL)
Figure 7. PV diagrams of a latex balloon.

40

Figure 7 shows the PV diagram of the balloon. The PV diagram was measured by injecting water into the balloon
with syringes (SS-50ESZ and SS-10ESZ, Terumo Corporation, Japan). With the balloon volume up to 12 mL, theinjection volume step was 12 mL. After that, the water was injected with the step of 1 mL. The internal pressureof balloon was measured with a pressure gauge (GC-31-174, NAGANO KEIKI CO., LTD., Japan) at each balloon
volume. The peak internal pressure on the PV diagram was 8 kPa when the volume of the balloon was around 5
mL. Beyond this, the internal pressure decreased as the volume increased. It was confirmed that the balloon
deformed greatly at a threshold value of about 8 kPa and higher.

FABRICATION AND EVALUATION OF THE ENTIRE SYSTEM
Fabrication
Figure 8 shows the fabricated system. The system consists of a positive displacement pump (V-15, Iwaki
Corporation, Japan), the fluidic switching valve developed in this study, syringes (SS-50ESZ, Terumo
Corporation, Japan), tubing connectors (TPX Tubing Connector 3549, Sanwaplatec Co., Ltd., Japan), balloons(polka dot balloon, Tiger Rubber Co., Ltd., Japan) and throttle valves (KT-6, Cole-Parmer LLC, USA). The
balloons are connected to the fluidic switching valve via the tubing connector. We employed water as the working
fluid in the following experiments.

SyringeTubing connector

BalloonPump

Fluidic Switching valve
Figure 8. Actual view of the entire system.

Throttle valve

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Evaluation
Just as mentioned in the previous section, the flow rate of main flow was varied from 0 to 26.7 mL/s with 20 equalintervals. Also, the cross-section ratio of the throttle valve was from 59% to 100% with 5 intervals. The flow rateof control flow 1 was set to 7.8 mL/s, and the control flow 2 was closed. The main flow and control flow 1 wasintroduced at t = 0 s, and after a sufficient time (8 s: time for the syringe to be empty), the flow was stopped. The
deformation of the balloon was visually recorded with the different combination of parameters of the main flowrate and the cross-section ratio of the throttle valve.
Figure 9 summarizes the experimental result. We categorized the behavior of the balloon as follows: (a) both of
the balloons are always inflated with or without control flow, (b) any of the balloon inflates but not deflates when
the control flow is stopped, (c) any balloon inflates and deflates when the control flow is stopped (desired
behavior), (d) one or both of the balloons did not inflate much, and (e) both of the balloons did not inflate at all.
Figure 10 shows the deformation of the balloons at a cross-section ratio of 59% and a main flow rate of 9.5 mL/s.
The time required for balloon deflation under this condition was 4 s for the blue balloon and 6 s for the yellow
balloon. Figure 11 shows the deformation of the balloon at a cross-section ratio of 59% and a main flow rate of
10.8 mL/s. Under this condition, the time required for balloon deflation was 12 s for the blue balloon and 83 s for

Cross-section ratio

90%(a) Expand both balloons(b) Expand too much

70%(c) Desirable deformation(d) Not expand enough(e) Not expand

50%
0

10
20
Volume flow of main flow (mL/s)
Figure 9. The experimental result.

(a) t=0 s(b) t=8 s(c) t=12 s(d) t’=0 s(e) t’=8 s(f) t’=14 sFigure 10. Balloon’s deformation under cross area: 59%, main flow: 9.5 mL/s.
(a) t=0 s: inflow started, (b) t=8 s: control flow 1 was stopped, (c) t=12 s: balloon contracted,(d) t’=0 s: inflow started, (e) t’=8 s: control flow 2 was stopped, (f) t’=14 s: balloon contracted.

(a) t=0 s(b) t=8 s(c) t=20 s(d) t’=0 s(e) t’=8 s(f) t’=83 sFigure 11. Balloon’s deformation under cross area: 59%, main flow: 10.8 mL/s.
(a) t=0 s: inflow started, (b) t=8 s: control flow 1 was stopped, (c) t=20 s: balloon contracted,(d) t’=0 s: inflow started, (e) t’=8 s: control flow 2 was stopped, (f) t’=83 s: balloon contracted.

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the yellow balloon. In Figures 10 and 11, (a/d) shows the initial state of the balloons at t = 0. When the water insyringe is fully injected at t = 8s, the balloons appear as shown in (b/e). After a sufficient time for deflection, theballoons restore their initial appearance as shown in (c/f). Time required for the balloons to be fully deflected
depends on the flow experimental condition. Note that (a-c) are for the blue balloon to be expanded, and (d-f) arefor the yellow balloon to be. These experimental results indicate that balloons can be selectively driven by setting
appropriate inflow conditions.
Here we discuss why the appropriate flow conditions (c) are distributed diagonally in Figure 9, and why there isa difference in time required to restore the initial appearance among the different conditions as shown in Figures
10 and 11. As the main flow rate increases, the inner pressure of the balloon increases because the velocity
difference between upstream and downstream of the balloon increases. Since the balloons required the thresholdpressure, 8 kPa, to largely deform, the flow rate of the main flow should be high enough to produce this thresholdpressure or higher. In other words, if the main flow rate is too low, the inner pressure of the balloon will not behigh enough to inflate the balloon sufficiently. To avoid this, the cross-section area ratio of the throttle valve
should increase to obtain the desired deformation. Therefore, the plots of desired deformation locate diagonallyin Figure 9.
When the flow rate of main flow is high and the balloon is inflating ((a) to (b), and (d) to (e) in Figures 10 and11), the inner pressure of the balloon once increases and then decreases by Δ!!"# according to the deformation of
the balloon, due to the nonlinear PV diagram shown in Figure 7. The smaller the Δ!!"# is, the easier it is for theballoon to deflate ((b) to (c), and (e) to (f) in Figures 10 and 11) when the control flow is stopped. In other words,
the deformability of the balloon depends on Δ!!"# , which depends on the deformation of the balloon when thecontrol flow is stopped. Then, the deformation of the balloon depends on the main flow rate. Therefore, it isconsidered that the time required for balloon’s deflation became longer as the main flow rate increased.

CONCLUTION
In this study, in order to simply realize the multi-degree-of-freedom motion of a soft robot, we proposed a method
to selectively inflate a balloon using a valve that can be operated by control flow input. The developed valve iscapable of generating a flow difference of approximately 0.25 mL/s between two outlets, and selective inflationof balloons is realized with a simple configuration for balloons having peaks in their PV diagram. This shows thepossibility of a new driving method for soft actuators, and shows that a multi-degree-of-freedom soft robot can be
driven only by fluid elements.

REFERENCES
1. NISHIOKA, Y., Pneumatic Soft Robot Arm, Journal of the Robotics Society of Japan, Vol.33, No.9, 2015, p.664-667.
2. KIM, S., LASCHI, C. and TRIMMER, B., “Soft robotics: a bioinspired evolution in robotics”, Trends in Biotechnology,
Vol.31, No.5, p.287-294, 2013.
3. SHEPHERD, F. R., ILIEVSKI, F., CHOI, W., MORIN, A. S., Stokes, A. A., MAZZEO, D. A., CHEN, X., WANG, M.
and WHITESIDES, M. G., Multigait Soft Robot, Proceedings of the National Academy of Sciences, Vol.108, No.51, 2011.
4. KATZSCHMANN, R., MARCHESE, A., RUS, D., Exploration of underwater life with an acoustically controlled soft
robotic fish, Science Robotics, Vol.3, No.16, 2018.
5. TAKEMURA, K., Soft Robotics using Electro-Conjugate Fluid (in Japanese), Measurement and control, Vol.57, No.11,
p.775-779, 2018.
6. CACUCIOLO, V., SHINTAKE, J., KUWAJIMA, Y., MAEDA, S., FLOREANO, D. and SHEA, H., Stretchable pumps
for soft machines, Nature, Vol.572, p.1-4, 2019.
7. GORISSEN, B., MILANA, B., BAEYENS, A., BROEDERS, E., CHRISTIAENS, J., COLLIN, K., REYNAERTS, D.
and DE VOLDER, M., Hardware Sequencing of Inflatable Nonlinear Actuators for Autonomous Soft Robots, Advanced
Materials, Vol.31, No.3, 2019.
8. ZATOPA, A., WALKER, S., and MENGUC, Y., Fully Soft 3D-Printed Electroactive Fluidic Valve for Soft Hydraulic
Robots, Soft Robotics, Vol.5, No.3, p.258-271, 2018.
9. WEHNER, M., TRUBY, R., FITZGETALD, D., MOSADEGH, B., WHITESIDE, G., LEWIS, J., WOOD, R., An
integrated design and fabrication strategy for entirely soft, autonomous robots, Nature, Vol.536, No.7617, p.451-455, 2016,
10. GALLOWAY, K., BECKER, K., PHILLIPS, B., KIRBY, J., LICHT, S., TCHERNOV, D., WOOD, R., GRUBER, D.,Soft Robotic Grippers for Biological Sampling on Deep Reefs, Soft Robotics, Vol.3, No.1, p.23-33, 2016.
11. POLYGERINOS, P., WANG, Z., GALLOWAY, K., WOOD, R., WALSH, C., Soft robotic glove for combined assistanceand at-home rehabilitation, Robotics and Autonomous System, Vol.73, p.135-143, 2015.
12. TESAŘ, V., Fluidic Valve for Reactor Regeneration Flow Switching, Chemical Engineering Research and Design, Vol.82,
No.3, p.398-408, 2004.

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Comparison of Model-Free Adaptive Displacement Control andModel Predictive Displacement Control
for Tap-Water-Driven Muscle
Considering Load Variation during Experiments
Satoshi TSURUHARA* and Ryo INADA** and Kazuhisa ITO****

Graduate School of Mechanical Engineering, Shibaura Institute of Technology
307 Fukasaku, Minuma-ku, Saitama, 337-8570 Japan(E-mail: md21074@shibaura-it.ac.jp)**

MEng, Graduate School of Mechanical Engineering, Shibaura Institute of Technology***

Department of Machinery and Control Systems, Shibaura Institute of Technology
307 Fukasaku, Minuma-ku, Saitama, 337-8570 Japan(E-mail: kazu-ito@shibaura-it.ac.jp)

Abstract. The muscle is well known to have strong asymmetric hysteresis characteristics and its characteristics
depend on the load applied to the muscle. This makes it difficult to design a controller for high-performance
displacement control considering load variations. In a previous study, model predictive control with a
servomechanism combining an inverse optimization algorithm and adaptive model matching, and model-free
adaptive control were applied to muscles, and high tracking control performance was achieved in the presence ofstatic modeling errors due to load variations in both control methods. However, the results for the load variation
during the experiment have not been discussed yet. Therefore, by comparing the two control methods, therobustness to load variations during the experiment was evaluated. The experimental results show that bothmethods have high tracking control performance equivalent to 0.013 mm in steady-state response after load
variations, and that MFAC is more robust than MPCS with AMM.
Keywords: Model-free adaptive control, Data-driven control, Model predictive control,
McKibben-type artificial muscle, Aqua-drive system

1. INTORODUCTION
The McKibben-type artificial muscle is widely researched in biorobotic engineering, medical engineering, andindustrial applications, owing to its many advantages including high flexibility, high power density, low weight,
low cost, and simple structure [1]. In particular, tap-water-driven muscles do not require any special power sourcesuch as compressor and pump and can be directly actuated by typical water supply network [2][3]. In addition,
they have advantages such as no fire hazards and a small burden on the environment, because tap-water is usedas the working fluid instead of pneumatic. Therefore, the tap-water-driven muscles are expected to be used in theabove-mentioned application that requires a high degree of cleanliness. However, the muscle is well known tohave strong asymmetric hysteresis characteristics caused by nonlinear contraction behavior and friction between
components. In addition, its characteristics depend on the load applied to the muscle. Therefore, it is difficult toderive a precise muscle model describing its behavior and to design a controller for high-performance
displacement control.
The purpose of this study is to propose a control method to solve this problem and to evaluate and examine its
robustness through experiments. In a previous study, model predictive control with servomechanism combining
inverse optimization and adaptive model matching (MPCS with AMM) is proposed, and high tracking control
performance was achieved even in the presence of static modeling errors for load variation applied to the muscle[4]. However, the method requires a complex mathematical model using an asymmetric Bouc-Wen model, which
is a quite time-consuming to be derived, identified, and evaluated. Moreover, in the presence of modeling errors,
the stability of the control system of the conventional method cannot be guaranteed. To solve these problems,model-free adaptive control (MFAC) [5], which does not require prior information about the mathematical modelof the control target and controller structure, was introduced, and high tracking control performance was achieved
while reducing the modeling effort [6]. In [4] and [6], the robustness of each control method to initial load variation
was confirmed, but their robustness to load variation during the experiments has not yet been discussed. Therefore,
in this paper, the robustness to load variation during the experiments is evaluated and compared for MFAC andMPCS with AMM.

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The rest of this paper is organized as follows. First, the concept and structure of the MPCS with AMM is described.
Next, introducing a full form dynamic linearization (FFDL) data model, the derivation of the MFAC controller
and the controller structure are described. Finally, experimental results of the proposed and conventional
controllers for load variation are presented to evaluate their effectiveness and robustness.

2. Model Predictive Control with Servomechanism by an inverse optimization
algorithm with adaptive model matching (MPCS with AMM)
In this section, a previous study, the model predictive control system with servomechanism by inverse
optimization with adaptive model matching, is briefly described using the block diagram in Figure 1 [4]. The
system was designed to provide robust control over the loads applied to the muscle. In the green region, the error
is calculated from the muscle system, and model which is a highly accurate mathematical model using theasymmetric Bouc-Wen model, and the adaptive algorithm is updated. Then, adaptive linearization and adaptivemodel matching are performed using the obtained estimated parameters to control the linearized muscle model tomatch the predictor output of the MPCS controller. On the other hand, in the blue region, the MPCS controller isdesigned by carrying out inverse optimization to achieve the pre-specified desired poles. Therefore, the control
performance is expected to be robust against load variations, and the robustness against static load variations was
confirmed.

Reference

MPCS
Controller+

Valve input voltage

Compensator（Linearization&
AMM）

McKibbenmusclesystem

Displacement

+

Error

McKibbenmusclemodel

Estimated
Parameter
Adaptive
Algorithm

Adaptive linearization and adaptive model matching system

Model predictive controller with servomechanism(predictor is prespecified and weights are obtained by the inverse optimization)

Figure 1. Block diagram of MPCS with AMM

3. Model-Free Adaptive Control (MFAC)
In this section, the MFAC system design is presented. The control system has many advantages, such as its
applicability to nonlinear discrete-time systems and its ability to guarantee rigorously asymptotic stability without
Lyapunov stability theory or key-technical lemma. After explaining the dynamic linearized data model, a pair ofthe estimation and control law for MFAC is derived.
3.1 Introduction to Dynamic Linearization Data ModelIn this section, the dynamic linearization data model, which is a fully equivalent to the unknown nonlinear
discrete-time system, is introduced to design the MFAC controller. In particular, the FFDL data model, which isone of the most common dynamic linearization models is explained. First, the unknown scalar nonlinear discretetime system is defined as follows.
𝑦(𝑘 + 1) = 𝑓 (𝑦(𝑘), 𝑦(𝑘 − 1), ⋯ , 𝑦(𝑘 − 𝑛𝑦 ), 𝑢(𝑘), 𝑢(𝑘 − 1), ⋯ , 𝑢(𝑘 − 𝑛𝑢 ))(1)

where 𝑢(𝑘), 𝑦(𝑘) ∈ 𝑅 represent the control input and output of the system for time 𝑘 , respectively, and𝑓(⋅): 𝑅𝑛𝑢 +𝑛𝑦 → 𝑅 is an unknown nonlinear function. Also, 𝑛𝑦 , 𝑛𝑢 ∈ 𝑍+ are two unknown orders of output andinput, respectively. The system represented by equation (1) can be rewritten as a dynamic linearization data model(in particular, the FFDL data model) by the following theorem. Here, in order to introduce the theorem, thenonlinear discrete-time system needs to satisfy the following two assumptions.

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Assumption 1
The partial derivatives of 𝑓(⋅) with respect to all variables are continuous.

□

Assumption 2
The nonlinear discrete-time system (1) satisfies generalized Lipschitz condition: There exists a real constant𝑏 ≥ 0 such that|𝑦(𝑘1 + 1) − 𝑦(𝑘2 + 1)| ≤ 𝑏 ‖𝑯𝐿𝑦+𝐿𝑢 (𝑘1 ) − 𝑯𝐿𝑦 +𝐿𝑢 (𝑘2 )‖(2)𝑇for all 𝑘1 , 𝑘2 ≥ 0. where 𝐻𝐿𝑦,𝐿𝑢 (𝑘) = [𝑦(𝑘), ⋯ , 𝑦(𝑘 − 𝐿𝑦 + 1), 𝑢(𝑘), ⋯ 𝑢(𝑘 − 𝐿𝑢 + 1)] ∈ 𝑅𝐿𝑦 +𝐿𝑢 is a vector
consisting of all outputs and inputs of the system up to pseudo-order 𝐿𝑢 (1 ≤ 𝐿𝑢 ≤ 𝑛𝑢 ), 𝐿𝑦 (1 ≤ 𝐿𝑦 ≤ 𝑛𝑦 ), whichis smaller than the order of system (1) such as 𝑛𝑢 , 𝑛𝑦 . The theorem for the FFDL data model is shown as follows.
□
Theorem (FFDL data model)
Assume that system (1) satisfies Assumption 1 and 2 given above. Moreover, assume that the norm of the change
in vector 𝑯𝑳𝒚,𝑳𝒖 (𝑘) , which consists of the changes in the inputs and outputs of the system, is nonzero, this is‖Δ𝑯𝑳𝒚,𝑳𝒖 (𝑘 )‖ ≠ 0 where Δ𝑯𝑳𝒚,𝑳𝒖 (𝑘) ≜ 𝑯𝑳𝒚,𝑳𝒖 (𝑘) − 𝑯𝑳𝒚 ,𝑳𝒖 (𝑘 − 1). Then, there exists a time-varying pseudogradient vector (PG vector) 𝝓𝒇,𝑳𝒚,𝑳𝒖 (𝑘) ∈ 𝑅𝐿𝑦+𝐿𝑢 such that the following equation (3) is satisfied.
Δ𝑦(𝑘 + 1) ≜ 𝑦(𝑘 + 1) − 𝑦(𝑘 ) = 𝝓𝑻𝒇,𝑳𝒚,𝑳𝒖 (𝑘)Δ𝑯𝒇,𝑳𝒚,𝑳𝒖 (𝑘)(3)𝑇where 𝝓𝒇,𝑳𝒚,𝑳𝒖 (𝑘) ≜ [𝜙1 (𝑘), ⋯ , 𝜙𝐿𝑦 (𝑘), 𝜙𝐿𝑦+1 (𝑘), ⋯ , 𝜙𝐿𝑦+𝐿𝑢 (𝑘)] . Furthermore, from Assumption 2 andequation (3), we can say that ‖𝝓𝒇,𝑳𝒚 ,𝑳𝒖 (𝑘)‖ ≤ 𝑏, which implies that the PG vector is bounded.

□

In the following sections, the FFDL data model derived by the above theorem is used to design the control
system.
3.2 Design the model-free adaptive controller based on the FFDL data modelFirst, it is necessary to estimate the PG vector to construct the FFDL data model at each time step, because thePG vector cannot be obtained analytically. For this purpose, the following evaluation function is introduced asfollows.
2

2

̂ 𝑓,𝐿 ,𝐿 (𝑘)) = |Δ𝑦(𝑘) − 𝝓̂ 𝑇𝑓,𝐿 ,𝐿 (𝑘)Δ𝑯𝐿 ,𝐿 (𝑘 − 1)| + 𝜇 ‖𝝓
̂ 𝑓,𝐿 ,𝐿 (k) − 𝝓̂ 𝑓,𝐿 ,𝐿 (k − 1)‖
𝐽𝑒 (𝝓
𝑦 𝑢
𝑦 𝑢
𝑦 𝑢
𝑦 𝑢
𝑦 𝑢

(4)

The evaluation function expressed by equation (4) represents, in the first term, the difference between the change
in output and the change in output represented by the FFDL data model using the estimated PG vector, and in thesecond term, the change in the PG vector with the weighting factor 𝜇 > 0. Then, minimizing the evaluation
function (4) with respect to the estimated PG vector, the following update law can be derived.

̂ 𝑓,𝐿 ,𝐿 (𝑘) = 𝝓
̂ 𝑓,𝐿 ,𝐿 (𝑘 − 1) +
𝝓
𝑦 𝑢
𝑦 𝑢

̂ 𝑇𝑓,𝐿 ,𝐿 (𝑘)Δ𝑯𝐿 ,𝐿 (𝑘 − 1))
𝜂Δ𝑯𝐿𝑦 ,𝐿𝑢 (𝑘 − 1) (Δ𝑦(𝑘) − 𝝓
𝑦 𝑢
𝑦 𝑢
2

(5)

𝜇 + ‖Δ𝑯𝐿𝑦 ,𝐿𝑢 (𝑘 − 1)‖

where 𝜂 ∈ (0,2] represents the step factor of the PG vector modification terms.
On the other hand, to derive the control law for MFAC, we consider the following 1-step ahead evaluation
function of control input as follows.
𝐽𝑐 (𝑢(𝑘)) = |𝑦𝑑 (𝑘 + 1) − 𝑦(𝑘 + 1)|2 + 𝜆|𝑢(𝑘) − 𝑢(𝑘 − 1)|2

(6)

where 𝑦𝑑 (𝑘 + 1) represents the reference signal and 𝜆 > 0 represents the weighting factor for the change in the
input. The evaluation function (6) represents in the first term, the difference between the reference and output of
the system, and in the second term, the change in the input. Substituting FFDL data model (2) into this evaluation

3

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function (6), and minimizing (6) with respect to the control input signal 𝑢(𝑘), and by replacing the PG vector withan estimate, the following control law can be derived as follows.

Δ𝑢(𝑘) =

̂ 𝐿 +1 (𝑘)(𝑦𝑑 (𝑘 + 1) − 𝑦(𝑘))𝜌𝐿𝑦+1 (𝑘)𝝓𝑦2

̂ 𝐿 +1 (𝑘)‖𝜆 + ‖𝝓𝑦−̂ 𝐿 +1 (𝑘) (∑𝐿𝑦 𝜌𝑖 𝝓
̂ 𝑖 Δ𝑦(𝑘 − 1 − 𝑖 ))
𝝓
𝑦
𝑖=1

(7)

2

̂ 𝐿 +1 (𝑘)‖𝜆 + ‖𝝓𝑦where 𝜌 ∈ (0,1], 𝑖 = 1, ⋯ , 𝐿𝑦 + 𝐿𝑢 represents the step factors for each PG vector element to make the controller
̂ 𝑖 represents the 𝑖-th element of the PG vector.
algorithm more flexible., and 𝝓
The block diagram of the above algorithm is shown in Figure 2. The MFAC performs the update law in equation(5) and the control law in equation (7) in online using only the change in input and output of the system. This
allows us to control nonlinear systems without explicitly using the mathematical models.

Figure 2. Block diagram of Model-Free Adaptive Control

4. Experimental results
In this section, the experimental conditions are described and a comparison between MFAC and MPCS with
AMM based on the experiment results is presented in order to evaluate the robustness to load variations duringexperiments.
4.1 Experimental conditions
The experimental conditions for MFAC and MPCS with AMM are shown in Table 1 and 2, respectively, andswitching the load applied to the muscle from 44 N to 68 N when experimental time is about 60 s to examine theload variations during the experiments. In additions, the experiment is conducted using the experimental circuit
shown in Figure 3 and reference signal is set step signal for 50 mm.
Table 1. Experimental conditions for MFAC
Value
0.01
1
1
1
[1 0.05]T

Unit
s
⋅
⋅
⋅
⋅

Table 2. Experimental conditions for MPCS with AMM
Value
0.1
5
5
0.99
0.95
1000𝐼16{0.6 0.7 0.8}

Unit
s
step
step
⋅
⋅
⋅
⋅

Item
Sampling time 𝑇𝑠
Weighting factors 𝜇, 𝜆Step factors 𝜌, 𝜂
Order of PG vector 𝐿𝑦 , 𝐿𝑢
̂ 𝑓,𝐿 ,𝐿 (1)Initial value of PG vector 𝝓
𝑦 𝑢

Item
Sampling time 𝑇𝑠
Predictive horizon 𝐻𝑝
Control horizon 𝐻𝑢
Coefficient of forgetting parameter 𝜆0
Initial value of forgetting parameter 𝜆(0)Covariance matrix 𝑃(0)Desired pole4

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Figure 3. Experimental circuit

4.2 Comparison of MFAC and MPCS with AMM by Experiments
Based on the above conditions, Figure 4 shows the comparison results of the control performance of MFAC andMPCS with AMM considering the load variation during the experiments. From this figure, it can be confirmedthat although the control performance deteriorates around 60 s for the load change, high tracking control
performance is achieved for both control methods in the steady-state response after 100 s. In addition, this figure
and Table 3 show that the tracking control performance of MFAC and MPCS with AMM is found to be almost
equal in the steady-state response. These results imply control performance that is generally close to theperformance limit of the sensor is obtained because the adaptive systems are introduced in both control methods.
On the other hand, from the evaluation of the maximum absolute error of Table 3, the control performance of
MFAC in the transient response immediately after load variation is better that that of the MPCS with AMM.

Load change

44 N

68 N

Figure 4. Comparison MPCS with AMM and MFAC for load variations during experiments
Table 3. Evaluation results of the control performance for Figure 4
MFAC
MPCS with AMM
0.013
0.013
7.85
15.2

Evaluation method and interval
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Figure 5 shows the results of tracking error in the steady-state response using 13 samples of data, (a) the transient
response after the load variation, (b) the steady-state response before (35 ~ 55s) and after (100 ~ 120s) the load
variation. Figure 5 (a), it can be seen that there is not much difference between the control performance of MFAC
and MPCS in the transient response immediately after the load is changed. On the other hand, the results after
load variation of Figure 5 (b) shows that the MFAC and the MPCS with AMM have almost the same control
performance in terms of the minimum value, however, the MFAC has a narrower quartile range and a narrower
maximum-minimum range than the MPCS with AMM. In addition, comparing the control performance before
and after the load variation, it can be seen that MPCS with AMM has higher control performance before the load
variation, while MFAC has higher control performance after the load variation. Therefore, these results show that
the MFAC is clearly more robust to load variation than the MPCS with AMM. The reason for this is that theseresults implies that the MFAC has a higher ability to sustain the adaptive capability of the PG vector. Moreover,
MFAC is a reasonable result because of its guaranteed asymptotic stability. On the other hand, although MPCS
with AMM can sometimes achieve high tracking control performance, it cannot reduce the modeling error due tothe lack of persistent adaptive capability, resulting in bad control performance in some cases. Furthermore, MPCS
with AMM is a reasonable result because stability is not guaranteed in the presence of modeling errors.

MFAC

MPCS with AMM

MFAC

MPCS with AMM(a) The transient response(b) The steady-state response (before and after load variation)
Figure 5. Box plots of experimental results for load variations during experiments (13 samples each)

5. Conclusion and Future work
In this paper, model predictive control with servomechanism combining inverse optimization and adaptive model
matching, one of the model-based control, and model-free adaptive control, one of the data-driven control, areapplied to tap-water-driven muscle, and the control performance and robustness against load variations duringexperiments are described. First, the previous study, MPCS with AMM is briefly described. Then, the full form
dynamic linearization data model, which is different from the mathematical model, is introduced and describes
the design of a model-free adaptive control system based on this model. The experimental results show that both
MPCS with AMM and MFAC have high tracking control performance of about 0.013 mm in steady-state response
after load variations, but MFAC is more robust than MPCS with AMM even though no exact mathematical modelis used. In the future, we will propose a method for adjusting parameters and initial values based on an
optimization algorithm.

REFERENCES
1. Andrikopoulos, G., Nikolalopoulos, G., Manesis, S., A Survey on Applications of Pneumatic Artificial Muscles,
Proceedings of 19th Mediterranean Conference on Control and Automation, 2011, pp. 1439-1446.
2. Conrad, F., Trends in Design of Water Hydraulics – Motion Control and Open-Ended Solutions, Proceedings of the 6th
JFPS International Symposium of Fluid Power, 2005, pp. 420-431.
3. Miyakawa, S., Aqua Drive System: A Technology Using Tap Water and Its Applications, Proceedings of the 8 th JFPSInternational Symposium on Fluid Power, 2011, pp. 26-37.
4. Tsuruhara, S., Inada, R., Ito, K., Model Predictive Displacement Control Tuning for Tap-Water-Driven Muscle by Inverse
Optimization with Adaptive Model Matching, Proceedings of the 10 th International Conference on Fluid Power
Transmission and Control -ICFP2021, 2021, pp. 286-292.
5. Hou, Z., Xiong, S., On Model-Free Adaptive Control and Its Stability Analysis, IEEE Transactions on Automatic Control,Vol. 64, No. 11, 2019, pp. 4555-4569.
6. Tsuruhara, S., Ito, K., Design of Model-Free Adaptive Displacement Control System for Tap-Water-Driven Muscle,
Proceedings of the Conference on Japan Fluid Power System Society in 2021 Spring, 2021, (in Japanese, Submitted).

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Experimental study on Dual-Layer Type Vortex Cup Driven by Aqua
Drive System
Wataru KOBAYASHI*, Tetsuya AKAGI* and Shujiro DOHTA**

Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science
1-1 Ridai-cho, Kita-ku, Okayama city, Okayama, 700-0005 Japan(E-mail: kobayashi@are.ous.ac.jp)
Abstract. This paper concerns with experimental study on a vortex cup. The cup has a vortex chamber inside the cup
and it can generate a swirl flow causing negative pressure in the chamber. There exists problems of the cup to besolved: improvement of the suction force and application of other working fluids. This study applies aqua drive system
as a driving source to the cup to validate the performance of the cup. In addition, a modification of the shape of thecup is considered. The approach adopts a folding flow path inside the cup. Experiments of a prototype dual-layer cup
are carried out to validate the performance of the cup with change of designed parameters: a chamber height, a flow
path width and a gap height. As a result, it is confirmed that the cup with redesigned parameters can achieve four times
as large as that of the conventional cup.
Keywords: Vortex cup, Aqua drive system, Dual-layer, Swirl flow, Optimal design

INTRODUCTION
A vortex cup is a new non-contact type handling method by using vortex levitation [1]. Traditional methods toconvey semiconductor and glass substrates adopt a contact type handling method which use pneumatic end
effectors to hold them. This method, however, has some problems as follows: 1) it causes static electricity between
end effectors and substrates, 2) it has a potential risk to hurt substrates directly, 3) its handling force is inadequate
to convey large, heavy and extra-thin substrates [2].
To solve the problems described above, a Bernoulli-chuck type end effector shown in Figure 1 was developed.
The end effector can convey substrates with non-contact method. Compressed air is exhausted from bottom-side
nozzles and then exhausted air from clearance between the end effector and floor surface which ejects in a radial
direction discharges into the atmosphere [3]. This makes the center area of the end effector, which is in vortex
chamber, negative pressure. Thus, the end effector with the vortex chamber can hold substrates by negative
pressure, and exhausted air prevent the end effector from touching with floor surface. The end effector, however,requires relatively large flow rate and has small suction force [4].
On the other hand, some studies showed that swirl flow can also make negative pressure in the center of flow [5].
Then a new non-contact type end effector using swirl flow, which called vortex cup shown in Figure 2, wasdeveloped [6-10]. The air consumption of the vortex cup can be one fourth of that of the Bernoulli-chuck type end
effector. Thus, it indicates that the cup with swirl flow can solve the problems mentioned above. There exist
additional studies about swirl flow in the cup. Li [11] reported a flow analysis on the cup, especially in the vortex
chamber. It was shown that the center of swirl flow in the vortex camber moved and this phenomenon decreased
velocity of swirl flow. This also caused reduction of suction force of the cup. Kawakami [12] introduced
cylindrical/circular- truncated-cone shaped guides into the vortex chamber to prevent movement of the center ofswirl flow. It was confirmed that height and diameter of the guides affect to suction force of the cup. Bonaccorso[13] also installed an electric motor into the center of the vortex chamber and the motor can generate a swirl flow
without supplied air.

Figure 1. Bernoulli-chuck type end effector (XT661, SMC corporation).

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Figure 2. Schematic diagram of vortex cup using swirling flow.
This study concerns with a modification of the shape of the cup to improve the suction force of the cup and an
application of aqua drive system using tap-water as a working fluid. Although the conventional pneumatic cup isapplied to a conveyor of semiconductor as mentioned above, an aqua drive cup described in this paper can beapplied to a holding mechanism of underwater systems. Other driving sources without aqua drive system are
unsuitable in underwater environment because of their waterproof performance. On the other hand, if the aqua
drive cup is available, holding mechanism in underwater environment with no waterproofing can be realized. The
conventional vortex cup mentioned above has simple cylindrical vortex chamber and some improvements applied
inner guides in the chamber. On the other hand, the authors introduce a folding flow path inside a vortex cup andmake second-layer in order to increase the flow velocity and reduce the radius of a vortex chamber. The detail ofthe cup is described later. Experimental analysis of a prototype new-shaped cup is also conducted and effects ofdesigned parameters of the cup on the suction force are validated to show an optimal design of the new-shaped
cup. Moreover, additional experiments show the difference of driving sources: aqua drive and pneumatic systems.

AQUA DRIVE VORTEX CUP
Vortex cups mentioned above are generally driven by pneumatic systems to mount them on the non-contact type
handling devices such as end effectors. The cups can overcome the problems that the conventional handling
methods mentioned above have. Moreover, an expansion their application range to other driving sources isrelatively easy because the working principle of the cups is simple. The cups have no limitation of working fluids.
There are, however, few studies concerning with any other working fluids such as oil, water, and functional fluids.
This study focuses on water as working fluid of the cup and aims to show the difference between pneumatic andaqua drive cups.
First of all, an aqua drive system is applied to a vortex cup. The shape of the vortex cup is a cylinder with a simple
vortex chamber and an inlet port on side surface of the chamber. Figure 3 shows a tested aqua drive cup made by
3D printer (AFINIA H800+, Microboards Technology Inc). The diameter of the vortex chamber r and the heightof the vortex chamber h in Figure 3(b) are 70 mm and 40 mm, respectively. The inner diameter of inlet port is 8
mm. Inlet flow from the port runs along the inner surface of the vortex chamber and forms swirl flow.

(a) Schematic view of tested cup(b) Cross-sectional view of tested cup

Figure 3. Tested aqua drive vortex cup.
To validate its performance, which means a suction force of the cup, an experimental setup shown in Figure 4 isconstructed. The experimental setup consists of the tested cup, a jig to hold the cup, a force gauge (FGC-5,
NIDEC-SHIMPO CORPORATION) to measure the suction force of the cup, and a wire to connect the cup withthe force gauge. Its driving source is from tap-water in the author’s laboratory of which pressure is approximately
280 kPa: almost average pressure of tap-water in Japan. Experiments here are carried out by changing the diameterand the height of the cup to show the effect of them to the suction force of the cup. Measurements are carried out

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eight times and the average of them are shown in Figure 5. Additional experiments which change only driving
source from aqua drive to pneumatic system are also carried out and Figure 5 plots the results: blue lines and dotsshow the results of the aqua drive cup and green lines and dots show the results of the pneumatic cup. Note thatboth driving sources used same supplied pressure of 280 kPa.

Figure 4. Schematic diagram of experimental setup for suction force measurement.

(a) Suction force vs. camber height(b) Suction force vs. chamber diameter

Figure 5. Experimental results of single-layer type vortex cup.
The experimental results showed that the tested cup generates approximately 5 N of the suction force and little
difference exists when the diameter of the cup is changed from 40 to 60 mm and the height of the cup is changed
from 70 to 90 mm. In particular, the cup with the chamber diameter of 50 mm generates the maximum suction
force but the difference between the maximum and minimum suction forces is only 2.5 N.
On the other hand, the results also showed the difference of driving sources. It is confirmed that the suction forceof the aqua drive cup was larger than that of the pneumatic cup under all conditions in these experiments. In
particular, the suction force of the pneumatic cup was less than 4 N under all conditions. Notice here that the sizeof the cup was designed with consideration for application of the aqua drive system: the diameter of the inlet port,
the diameter and height of the vortex chamber. In other words, the tested cups were specialized in aqua drive
system and their dimension might be too large to apply to the pneumatic cup. The important point in theseexperiments is that we can apply other driving sources such as aqua drive system same as the conventional
pneumatic system. This suggests that application range of vortex cups can be expanded.

DUAL-LAYER TYPE VORTEX CUPThe foregoing experiments showed a possibility of a vortex cup driven by aqua drive system. The cup, however,
had insufficient suction force in practical. In addition, the cup constantly requires swirl flow in the vortex chamber

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to keep negative pressure when it works. Then, it is considered to modify the shape of the cup in order to improve
the performance of the aqua drive vortex cup. The shape of the conventional cup was simple cylinder and somestudies inserted cylindrical/circular- truncated-cone shaped guides as mentioned above. In this study, flow pathinside a vortex cup is folded to increase flow velocity of the swirl flow and also the suction force of the cup.
This approach was applied to the micro-bubble generator (Aqua air, idea-techno Co., Ltd.) shown in Figure 6.
The micro-bubble generator adopted a swirl flow way and it can generate bubbles with diameter of 1 - 60 μm.
The inlet flow runs along with inner side surface of the chamber of the generator and negative pressure occurs at
the center of the chamber. At the moment, micro-bubbles can be generated due to supplied gas to the center of thecamber. Thus, the operating principle of the generator is similar to that of the vortex cup. As seen in Figure 6, theflow velocity inside the generator can increase by folding the flow path at its turn-round points: it is shown asarrows in the figure.
On the other hand, the centrifugal force in the vortex chamber relates to swirl flow which also relates to the suction
force of the vortex cup. A centrifugal force F in general is expressed as Eq. (1) where m, ω, and r are a mass, anangular velocity and a radius, respectively. It is clear that a reduction of the radius and increase of the angularvelocity lead to gain of the centrifugal force. Thereby, if a radius of a vortex chamber is reduced by folding flow
path in a vortex cup as the micro-bubble generator did, the suction force of the cup can be gained without any
modifications of outer dimensions of the cup.

Figure 6. Operating principle of micro-bubble generator (Aqua air, idea-techno Co., Ltd.).
𝐹 = 𝑚𝜔2 𝑟.

(1)

A prototype of a vortex cup with a folding flow path is made to validate the effectiveness of the forementioned
approach. Figure 7 shows the comparison of conventional single-layer and proposed dual-layer type vortex cups.
As seen in the cross-sectional view of Figure 7, the structure of the right cup has dual-layer flow path inside thecup and parameters ri, ro, hc and hi exist. The parameters indicate an inner diameter of the vortex chamber, a widthof the flow path in a second-layer, a height of the flow path at the turn-around point, and a height of the vortex
chamber, respectively. Figure 8 shows the schematic view of the prototype dual-layer cup and its cross-sectionalview. The prototype cup realizes a dual-layer flow path inside the cup by assembling two parts made by 3D printermentioned above. Figure 9 shows the assembly parts of the prototype cup and the assembled structure enables theparameters ri, ro, hc and hi to be modified easily by replacing the parts although details of the modifications aredescribed below.

(a) Single-layer(b) Dual-layer

Figure 7. Comparison of single-layer and dual-layer type vortex cups.

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The 11th JFPS International Symposium on Fluid Power 2020(a) Schematic view of dual-layer cup(b) Cross-sectional view of dual-layer cup

Figure 8. Schematic diagram of tested dual-layer cup.

Figure 9. Assembly parts of tested dual-layer cup.

EXPERIMENTAL ANALYSIS OF DUAL-LAYER TYPE VORTEX CUPThe dual-layer cup has additional parameters such as a flow path width ro in the first-layer and a gap height hc,
that is, there exist four design parameters to optimize the performance in the suction force of the cup: a chamber
diameter ri, a chamber height hi, a flow path width ro and a gap height hc. Note that only the chamber diameter infollowing experiments is fixed so as not to change the vortex chamber area of the cup from that of the conventional
single-layer cup. In addition, other parameters such as a diameter of the inlet flow port, an outer diameter of thecup and a height of the cup are also same as the parameters of the single-layer cup mentioned in previous chapter
in order to compare the performance of the dual-layer cup with the single-layer cup. Experiments to measure thesuction force of the cup are carried out by using the experimental setup shown in Figure 4 and measurement
conditions and procedure are also same as them of the previous experiments. Table 1 shows the designed
parameters of a benchmark dual-layer cup of which outer dimension is almost equivalent to the one of the singlelayer cup.
Table 1. Designed parameters of benchmark dual-layer cup.
Parameters
Symbols
Value [mm]Flow path widthro10
Gap heighthc
10Chamber heighthi
50
Chamber diameterri
70
Inlet port diameter
8First experiment evaluates an effect of changing the flow path width ro on the suction force of the cup. Figure 10
shows the experimental results. The suction force was measured with changing of only the flow path width from
5 to 15 mm. The difference between aqua drive and pneumatic cups were also examined and the both results areplotted in Figure 10: blue lines and dots show the results of the aqua drive cup and green lines and dots show theresults of the pneumatic cup. As seen in Figure 10, the suction force of the aqua drive cup was drastically increased
due to the dual-layer structure. It is experimentally confirmed that the cup with the flow path width of 10 mm cangenerate the maximum suction force of approximately 22 N, which is almost three times larger than that of thesingle-layer cup. Nevertheless, the suction force of the pneumatic cup with dual-layer showed little change fromthat of the single-layer cup and also change of flow path width had little effect on the suction force.

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Figure 10. Experimental result on flow path width.
Second and third experiments evaluate effects of changing the gap height hc and the chamber height hi on thesuction force of the cup, respectively. Figures 11 and 12 show the experimental results. The suction force wasmeasured with changing of the gap height from 5 to 15 mm in the second experiment and with changing of thechamber height from 50 to 70 mm in the third experiment. Note that the second experiment used the redesigned
flow path width of 10 mm and the third experiment used the redesigned flow path width of 10 mm and gap heightof 15 mm taking account of the results of the first and second experiments. The difference between aqua drive
and pneumatic cups were also examined and the both results are plotted in Figures 11 and 12: blue lines and dotsshow the results of the aqua drive cup and green lines and dots show the results of the pneumatic cup.
The cup with the gap height of 5 mm generates the maximum suction force of approximately 22 N. The gap heightin the second experiment means the difference between the height of the first-layer and that of the second-layer
in the cup. The force difference in the experiment was only 1.7 N and this means change of the gap height had
little effect on the suction force of the cup. On the other hand, it is clear that the chamber height had strong
relationship with the suction force of the cup. The cup with the chamber height of 60 mm generates the maximum
suction force of approximately 25 N. There existed a large difference between the result with 60 mm chamberheight, which showed the maximum suction force, and that with 50 mm chamber height, which showed theminimum suction force. Nevertheless, the suction force of the pneumatic cup showed little change in bothexperiments. The size of the tested cups was designed with consideration for application of aqua drive system asmentioned above. In particular, the diameter of the inlet port was designed for application of aqua drive system.
Hence, the experimental results here showed same tendencies regardless of designed parameters and experimental
conditions.
Table 2 shows the parameters for the aqua drive dual-layer cup which mean the combination of the parameters
generating the maximum forces on each previous experiment. The maximum suction force of the redesigned cup,
which is approximately 25 N, can achieve four times as large as that of the conventional single-layer cup. Thus,
the modification with second-layer can increase the velocity of swirl flow in the chamber and then it can also
increase the suction force. To analyze the relationship between the designed parameters and the suction force indetail, theoretical consideration based on mathematical models of the cup and computational analysis are required
and it will be a next step of this study.

Figure 11. Experimental result on gap height.

Figure 12. Experimental result on chamber height.

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Table 2. Combination of designed parameters for dual-layer cup.
Parameters
Symbols
Value [mm]Flow path widthro15
Gap heighthc
10Chamber heighthi
60
Chamber diameterri
70
Inlet port diameter
8

CONCLUSIONSThis paper concerns with experimental study on a vortex cup which is a new non-contact type handling methodby using vortex levitation. The cup has a vortex chamber inside the cup and it can generate swirl flow. The
generated swirl flow makes negative pressure in the center of the vortex chamber. This is the working principleof the cup and considerable points of the cup are 1) improvement of the suction force of the cup, 2) application ofother working fluids such as oil, water and functional fluids.
This study applies aqua drive system as a driving source to the cup in order to validate the performance of the cup
compared with that of the conventional pneumatic driven cup. The experiment of the tested aqua drive cup madeby 3D printer was carried out and the relationship between the parameters of the cup and the suction force wasshown. Moreover, the additional experiments also showed the difference of two driving sources.
On the other hand, a modification of the shape of the cup was considered. The approach of the modification
adopted a folding flow path inside the cup, that is, there existed a second-layer on an outer side of a first-layer ofthe cup. The experiments of prototype dual-layer cup were carried out to validate the performance of the cup withchange of designed parameters: a chamber height hi, a flow path width ro and a gap height hc. As a result, it isconfirmed that the cup with redesigned parameters can achieve four times as large as that of the conventional
single-layer cup.
Future work of this study is to analyze the relationship between the designed parameters and the suction force indetail, especially for both aqua drive and pneumatic cups. Theoretical analyses based on mathematical models of
the cups and computational analyses will be conducted.

REFERENCES
1. BRANDT, E. H., Levitation in physics. Science, 1989, 243-4889, pp. 349-355.
2. MORISAWA, T., et al., A non-contact holder using airflow, JFPS International Journal of Fluid Power System, 2019, 113, pp. 104-109.
3. WALTHAM, C., BENDALL, S., KOTLICKI, A., Bernoulli levitation, American Journal of Physics, 2003, 71-2, pp. 176179.
4. DINI, G., FANTONI, G., FAILLI, F., Grasping leather plies by Bernoulli grippers, CIRP annals, 2009, 58-1, pp. 21-24.
5. LI, X., KAWASHIMA, K., KAGAWA, T., Analysis of vortex levitation, Experimental Thermal and Fluid Science,
2008, 32-8, pp. 1448-1454.
6. IIO, S., Research on a non-contact handling system using swirling flow (influence of cup-to-work gap thickness on swirling
flow characteristics), SICE Trans Ind Appl, 2008, 7-12, pp. 79-84.
7. FALLAH, A., et al., Numerical analysis of a non-contact surface adhesion system based on vortex cup for wall climbing
robots, 2013 First RSI/ISM International Conference on Robotics and Mechatronics (ICRoM). IEEE, 2013, pp.368-373.
8. ZHAO, J., LI, X., Effect of supply flow rate on performance of pneumatic non-contact gripper using vortex
flow, Experimental Thermal and Fluid Science, 2016, 79, pp. 91-100.
9. LI, X., KAWASHIMA, K., KAGAWA, T., Dynamic characteristics of vortex levitation, 2008 SICE Annual Conference.
IEEE, 2008. pp. 1175-1180.
10. WU, Q., YE, Q., MENG, G., Particle image velocimetry studies on the swirling flow structure in the vortex
gripper, Proceedings of the Institution of Mechanical Engineers, Part C: Journal of Mechanical Engineering Science, 2013,
227-9, pp. 1927-1937.
11. LI, X., et al., Computational fluid dynamics study of a noncontact handling device using air-swirling flow, Journal ofengineering mechanics, 2011, 137-6, pp. 400-409.
12. KAWAKAMI, C., NAKAO, M., NISHIMURA, Y., KONDO, E., Increased sucking force of Vortex Cup attached a
Cylinder in the Vortex Chamber, Proceedings of JFPS Autumn Conference, (2015), pp.56-58.
13. BONACCORSO, F., BRUNO, C., LONGO, D., MUSCATO, G., Structure and model identification of a vortex-based
suction cup, Advances in Mobile Robotics, 2008, pp. 303-310.

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Organized Session | Organized session

Soft Actuator 2[OS5-2-01] Pneumatic Source Proposal for Improving Portability andResponsiveness of Artificial Muscle via Dimethyl Ether Phase Change
and Combustion〇Ryuto Enjo1, Manabu Okui1, Taro Nakamura1 （1.Chuo university）[OS5-2-02] Development of Six-Legged Mobile Robot Using Tetrahedral Shaped
Pneumatic Soft Actuators〇Kenta Hase1, Tetsuya Akagi1, Shujiro Dohta1, Takashi Shinohara1, Wataru Kobayashi1, So
Shimooka2 （1.Okayama University of Science, 2.Okayama University）[OS5-2-03] Development of the Transfer System for Bedridden Elderly and Disabled
People using Pneumatic Actuators〇Feifei Cho1, Keisei Kato1, Ryota Endo1, Takumi Kobayashi2, Tetsuya Akagi2 （1.National
Institute of Technology Tsuyama College, 2.Okayama University of Science）[OS5-2-04] Development of Pneumatic Drive Pipe Inspection Robot using Radial
Bending Type Soft Actuator〇Takashi Shinohara1, Hikaru Furuya1, Tetsuya Akagi1, Shujiro Dohta1, Takumi Kobayashi1, So
Shimooka2 （1.Okayama University of Science, 2.Okayama University）©The Japan Fluid Power System Society

OS5-2-01

The 11th JFPS International Symposium on Fluid Power 2020

Pneumatic Source Proposal for Improving Portability andResponsiveness of Artificial Muscle via Dimethyl Ether Phase
Change and Combustion
Ryuto ENJO*,Manabu OKUI*,Taro NAKAMURA**Department of Precision Mechanics, Faculty of Science and Engineering
Chuo University
1-13-27 Kasuga, Bunkyo-ku, Tokyo, 112-8551 Japan(E-mail: r_enjo@bio.mech.chuo-u.ac.jp)
Abstract. Pneumatic drive systems are widely used in assistance devices and flexible drive robots owing to theirvarious characteristics. Compressed air generated by engine compressors and large motors drives these devices, but suchlarge and heavy air pressure sources limit the mobilization of pneumatic drive devices. Herein, we propose a two-stage
system using dimethyl ether (DME) phase change and combustion to develop the mobilization of pneumatic drive devices.
Also, the artificial muscle driven by DME combustion is introduced and prototyped to improve responsiveness. Besides, we
experimentally investigated DME combustion characteristics. The result shows a range of combustible air-DME fuel ratios
and an air-DME molar ratio of 0.07% that generated the highest pressure. Furthermore, we measured the response time anddisplacement of the artificial muscle when DME burned in the device. The results indicate that this device drives rapidly
than that of previous research.
Keywords: Air-pressure source, artificial muscle, combustion, pneumatic actuators, pneumatic systems

INTRODUCTION
Pneumatic drive systems are widely used in assistance devices [1] and flexible drive robots [2] due to theirlightweight, high output, and excellent flexibility characteristics. However, the problems of pneumatic drive
systems are air pressure sources and low responsiveness. Compressed air generated by engine compressors andlarge motors is mainly used to drive these devices, but such large and heavy air pressure sources are barriers tomobilizing pneumatic drive devices. Thus, the construction of a portable air pressure source can expand theapplication range of such devices.
Along these lines, many mobile pneumatic drive systems have been developed. Among them is an air pressuresource used by three points of dry ice [3] that can silently release gas at 0.42 MPa. However, at normal
temperatures, the energy required to produce compressed air cannot be safely stored because vaporization could
inadvertently occur at any time. Our previous research evaluated an air pressure drive system to develop a
portable air pressure source. However, no portability method could be identified [4] (Figure 1)．In addition toprevious methods, we developed a hybrid pneumatic source [5], which combines the phase changes of liquidcarbon dioxide and dimethyl ether (DME). The prototype is able to provide much more amount of compressedgas compared to existing method, but DME had problems, such as flammability and thermal drop problems.
Here, the responsiveness of pneumatic actuators is investigated. The compressed air must activate pneumatic
actuators, but the compressible of the air causing the response delay. Also, compressed air is generally supplied
through the narrow tube, which causes a response delay.
To improve the response of pneumatic actuators, pneumatic artificial muscles and Magneto-rheological (MR)fluid method has been developed [6]. In this method, elastic uses artificial air muscle, and instantaneous force isgenerated using MR fluid brake that can change the viscosity. Although this method can generate great force
instantly, it needs advanced preparation to drive the devices.
Therefore, the study aims to improve the portability and responsiveness of pneumatic drive systems. To do so,
we focus on the physical properties of flammable DME and consider obtaining energy by DME combustion.
Also, we consider obtaining energy in two stages via phase change and DME combustion. The proposed method,which uses the gas–liquid phase change of DME at a high flow rate per mass, does not require complicatedauxiliary systems and equipment.
Besides, we think that DME combustion can solve the problems of the hybrid pneumatic source, which is theflammability or thermal drop. To improve the responsiveness of the pneumatic drive system, we proposed thepneumatic system driven by DME combustion. This method can generate large instantaneous force using DME
combustion inside the pneumatic system. Herein, we proposed the concept of a two-stage pneumatic source andinvestigated DME combustion characteristics. Also, we proposed a pneumatic system driven by DME
combustion and investigated its responsiveness when air-DME mixed gas burned inside the artificial muscle.

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The 11th JFPS International Symposium on Fluid Power 2020

Generable pressure [MPa]1.2
Ideal Performance

1

Chemical reaction

0.8
Compressor measured value(0.4 MPa)

0.6
0.4

Compressor measured value(0.2 MPa)
DME

0.2

Compressor measured value(0.0 MPa)

Dry-ice
0
0.000

0.500

1.000

1.500

Generated flow per mass [NL/g]

Figure 1. Mobility of pneumatic source flowchart [4]TWO-STAGE DME USAGE SYSTEM PHASE CHANGE AND COMBUSTION
An overview of the proposed system is shown in Figure 2. Pressurization and decompression occur at regularintervals, and a leveraged pneumatic artificial muscle is used as a pneumatic actuator. The main components ofa two-stage usage system are explained below.
(a) Pneumatic actuators are driven via the phase change of DME(b) After driving the actuators, DME exhaust gas is collected and put into a container(c) Energy is obtained via combustion, where the DME is ignited using a spark plug in the container.
DME gas-liquid phase change can generate 0.4 MPa of pressure stably. DME does not release particulate
pollutants and smoke because it is oxygenated fuel and has no carbon bonds. Additionally, DME is a gas at
normal temperatures, and its saturated vapor pressure is low (= 0.6 MPa), so it can easily be stored as a liquid.
Based on these properties, DME can be suitably used in a two-stage system.
2 port valve

Artificial air muscle
DME
PC

Combustion
Artificial air muscle

Spark plug

Figure 2. Outline of two-stage application system of phase change and combustion of DME

COMBUSTION EXPERIMENT
Experimental Purpose
The experiment aims to verify the air-DME fuel ratio range for combustion via spark plugs and measure thepressure.

Experimental Method
Figure 3 shows the schematic representation and photo of the experimental device．A tank, pressure sensor(MPS-R33RC-NGAT, MYOUTOKU, Japan), regulator (RVUM6-6, PISCO, Japan), two-port valve (MVP62,PISCO, Japan), spark plug (Hirano giken, Japan) were used. A gas mixture of air-DME fuel was stored in thetank in advance and was injected into a semi-closed container for 3 s using a pressure gauge, a regulator, and a
two-port valve. The mixed gas was ignited via a spark plug 2 s after the injection. When the mixed gas burned,
the pressure was measured by changing the mixing ratio. The molar ratio parameter, which was set to 0.03%–
0.14%, was based on the theoretical air-DME ratio (0.067%). The experiment was conducted in a room
maintained at 25°C, and an atmospheric pressure of 0.1013 MPa was used in the calculations. Eachmeasurement was performed thrice. A 1.0 × 1.3-cm2 hole was created above the semi-closed container because
the container’s pressure rapidly rises when mixed gas is burned.

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The 11th JFPS International Symposium on Fluid Power 2020Hole(1.0 cm ×1.3 cm)

Hole (1.0 cm×1.3 cm)
combustion

Tank (DME&Air)

Spark plug
Pressure sensor
0

Figure 3. Image of experimental device

Experimental Results and Discussion
Figure 4 shows the pressure variation with the mole ratio parameter when mixed gas burned. The timescale inthis graph is unsynchronized. The error bars in Figure 4 represent the standard deviation. The result shows therange of combustible air-DME fuel ratios. The pressure increment occurred due to combustion in the
combustible range, and the air-DME molar ratio of 0.07% generated the highest pressure. Considering thetheoretical air-DME ratio (0.067%), the highest pressure was generated when air and DME reacted inappropriate proportions.
0.02

0.05%

0.07%
0.08%
0.09%

0.01

0.10%
0.11 % (No Peak)0.005

0.12 % (No Peak)0-0.005
0

0.02

0.04

0.06

Time [s]

0.08

0.1

Dischange Pressure [MPa]Dischange pressure [MPa]

0.06%

0.015

0.025
0.020

0.015
0.010
0.005

0.000
0.05

0.06

0.07

0.08

0.09

0.10

DME molor ratio [%]Figure 4. Pressure variation of mixed gas when burned

ARTIFICIAL MUSCLE DRIVEN BY DME COMBUSTION
Figure 5 shows the outline of the artificial muscle driven by DME combustion. The McKibben pneumatic
artificial muscle was used as the pneumatic system. The artificial muscle driven by DME combustion is
explained below．
① The mixed gas was injected into the artificial muscle② The gas mixture was ignited by the spark plug in the artificial muscle③ The rapid pressure increment drove the artificial muscle due to combustion
The method’s responsiveness was explained using a simple model (Figure 5). We assumed that the artificial
muscle (length and diameter of 100 and 12 mm, respectively) was used and sparks at its center. The distancefrom the spark point to the surface of the artificial muscle is 6 mm. The air propagates to the surface of theartificial muscle in 1.9 ms after the mixed gas combustion because DME propagates at a speed of 3.1 m/s [7].
Therefore, based on the above, this method effectively improves the responsiveness of pneumatic drive systems.

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Experiment
The pressure and displacement were measured when the air-DME mixed gas burned to confirm the proposed
method’s displacement responsiveness.
Figure 6 shows the experimental setup. Here, the McKibben pneumatic artificial muscle (length and diameter of
100 and 12 mm, respectively), which was used as the pneumatic artificial muscle, was filled with air-DME
mixed gas in advance. The mixture ratio of the mixed gas is 0.07%, which generated the highest pressure in theprevious experiment. One end of the artificial muscle is attached to a fixed end, whereas the other end to theslide rail, which moves freely along with the displacement of the artificial muscle. Sparks are generated inside
the artificial muscle, which ignites the mixed gas. When the mixed gas burned, the pressure was measured using
a pressure sensor (MPS-R33RC-NGAT, MYOUTOKU, Japan). Also, the artificial muscle’s displacement wasmeasured using a laser displacement meter (HG-C1200, Panasonic, Japan). Each measurement was performed
thrice.
Pressure sensor

Move freely
Sparker

0

100 mmMixed Gas

PC

Move freely

Pressure sensor

Sparker

0

PC

12 mm

Load cell

Load cell

Laser displacement meter

Laser displacement meter

Figure 5. Image of artificial muscle driven by DME combustion

Artificial muscle

Pressure sensor

Load celll = 100 [mm]Sparker

Laser displacement meter

Figure 6. Experimental setup for displacement and contraction force measurement

Experimental Results and Discussion
Figure 7 shows the responsiveness of pressure and displacement of the artificial muscle in DME combustion.
The horizontal axis represents the time from the spark, the first vertical axis represents the artificial muscle’s
internal pressure, and the second vertical axis represents the artificial muscle’s displacement. The delay time ofdisplacement was 0.028 s and the reference value of the time constant was 0.037 s (different from the original
meaning). Therefore, these results indicate that this system improves responsiveness compared to theconventional pneumatic drive system.
From Figure 6, it takes 0.02 s for the artificial muscle to shrink after the pressure increases. It can also be
confirmed that the artificial muscle’s displacement oscillates like a damped wave. Besides, it is thought that the
McKibben artificial muscle has an elastic component because it is made of a rubber tube.

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Pressure [MPa]

0.2

150

Pressure145

Displacement

0.15

140
0.1

135
0.05

130

0

0

0.05

0.1

0.15

0.2

0.25

0.3

Displacement [mm]

155

0.25

125
0.35

Time [s]Figure 7. Pressure and displacement of artificial muscle in DME combustion

EXPERIMENT ON CONTRACTION FORCE RESPONSIVENESS OF DMECOMBUSTION
The pressure and force were measured when the air-DME mixed gas burned to confirm the proposed method’s
force responsiveness.

Experimental Method
Figure 6 shows the experimental setup. Just as in the displacement experiment, the McKibben pneumatic
artificial muscle (length and diameter of 100 and 12 mm, respectively) was used here and was filled with the airDME mixed gas in advance. The mixture ratio of the mixed gas is 0.07%, which generated the highest pressurein the previous experiment. To keep the artificial muscle’s length constant, both ends of the artificial muscle are
fixed. Sparks are generated inside the artificial muscle, which ignites the mixed gas. When the mixed gas burned,
the pressure was measured using a pressure sensor (MPS-R33RC-NGAT, MYOUTOKU, Japan), and theartificial muscle’s contraction force was measured using a load cell (LUX-B-ID, KYOWA, Japan).

Experimental Results and Discussion
Figure 8 shows the responsiveness of pressure and contraction force of the artificial muscle in DME combustion.
The horizontal axis represents the time from spark, the first vertical axis represents the artificial muscle’s
contraction force, and the second vertical axis represents the artificial muscle’s internal pressure. It was
confirmed that it takes 0.0057 and 0.0154 s for the pressure increment and the contraction force effect after
spark, respectively. Therefore, these results indicate that this system improves responsiveness, as well as thedisplacement responsiveness.
From Figure 8, it can be confirmed that there is a time lug caused between the increased pressure andcontraction force. It is thought that the McKibben artificial muscle has an elastic component. The maximum
contraction force was about 44 N. In the future, we will investigate different diameters and lengths of theartificial muscle to produce artificial muscle that can generate high output.

Force

Force [N]40

Pressure

0.08

30

0.06

20

0.04

10

0.02

Pressure [MPa]

0.1

50

0

0

0

0.05

0.1

0.15

0.2

0.25

Time [s]Figure 8. Pressure and force variation with time of artificial muscle in DME combustion

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CONCLUSIONS
Here, we proposed a two-stage system using DME phase change and combustion to improve the portability ofpneumatic drive systems. Also, we confirmed the range of the air-DME fuel ratio for combustion. Furthermore,
we proposed an artificial muscle driven by combustion to improve responsiveness. Besides, we measured the
responsiveness of displacement and contraction force when the air-DME mixed gas burned. Consequently, highresponsiveness (the delay time of displacement and contraction force were 0.028 and 0.015 s, respectively) was
achieved.
In the future, we will investigate the responsiveness driven by combustion by changing different diameters andlengths of the artificial muscle.

ACKNOWLEDGMENTS
Research supported by Yashima Environment Technology Foundation.

REFERENCES
1. SASAKI, D., NORITSUGU, T., TAKAIWA, M., Development of Active Support Splint Driven by Pneumatic Soft
Actuator (ASSIST), In Proceedings of the 2005 IEEE International Conference on Robotics and Automation, 2005,
pp.520-525.
2. WANG, Z., POLYGERINOS, P., OVERVELDE, J. T., GALLOWAY, K. C., BERTOLDI, K., WALSH, C. J.,
Interaction Forces of Soft Fiber Reinforced Bending Actuators, IEEE/ASME Transactions on Mechatronics, 2017, 22-2,
pp.717-727.
3. WU, H., KITAGAWA, A., TSUKAGOSHI, H., PARK, S. H., Development and Testing of a Novel Portable PneumaticPower Source Using Phase Transition at the Triple Point, Proceedings of the JFPS International Symposium on Fluid
Power, 2005, 223-6, pp.310-315.
4. OKUI, M., NAGURA, Y., ISHIIKAWA, S., YAMADA, Y., NAKAMURA, T., A Pneumatic Power Source Using a
Sodium Bicarbonate and Citric Acid Reaction with Pressure Booster for Use in Mobile Devices, In 2017 IEEE/RSJ
International Conference on Intelligent Robots and Systems (IROS), 2017, pp.1040-1045.
5. OKUI, M., NAGURA, Y., YAMADA, Y., NAKAMURA, T., “Hybrid Pneumatic Source Based on Evaluation of Air
Compression Methods for Portability,“ IEEE Robotics and Automation Letters, 2008, 3, pp.819-826
6. TOMORI, H., MAJIMA, T., ISHIHARA, H., NAKAMURA, T., “Throwing Motion with Instantaneous Force Using a
Variable Viscoelastic Joint Manipulator”, Journal of Intelligent Material Systems and Structures, 2017, 28, pp.999-1009 .
7. MOGI, T., OKADA, K., WADA, Y., HORIGUCHI, S.,YOSHIDA, M.,IKEDA, K.,NANBA, M., “Hazards of Large
Scale Pool Fire of Dimethyl Ether” Japan Society for safety Engineering, 2007, 46

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Development of Six-Legged Mobile Robot
Using Tetrahedral Shaped Pneumatic Soft Actuators
Kenta HASE*, Tetsuya AKAGI*, Shujiro DOHTA*, Takashi SHINOHARA**,
Wataru KOBAYASHI* and So SHIMOOKA****

Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science**
Science Dream lab., Okayama University of Science
1-1,Ridai-cho, Kita-ku, Okayama 700-0005, Japan(E-mail: akagi@are.ous.ac. jp)***
Graduate School of Natural Science and Technology, Okayama University
3-1-1, Tsushima-naka, Kita-ku, Okayama 700-8530, Japan

Abstract. Based on elderly society and decreasing birth rates in Japan, the rehabilitation device that can be used inhome or hospital without assistance has been desired. In the previous study, a tetrahedral shaped flexible actuator using
three extension type flexible pneumatic actuators was proposed and tested as a wrist rehabilitation device. The
tetrahedral shaped actuator can extend longitudinally and bend toward any radial direction. In this study, a mobile
robot using six tetrahedral shaped actuators that can use a core training and amusement for patients and elderly is
proposed and tested. In this paper, the construction and operating principle of the tested robot was described. The
control system using a micro-computer and 18 on/off valves was also described. In addition, the walking and rotating
tests using the tested robot was carried out.
Keywords: Tetrahedral shaped pneumatic soft actuator, Extension type flexible pneumatic actuator, Mobile robot,
Home-based compact 3D simulator

INTRODUCTION
In 2020, the ratio of the Japanese elderly became more than 26 % [1]. In 2040, according to the statisticalprediction, the ratio will be growing up to 36%. Under these circumstances, various assisting and rehabilitation
devices to help welfare work for the elderly and disabled using pneumatic soft actuators[2] have been actively
developed[3-10]. As a pneumatic power assisting device, the wearable power assist suit for nursing care using
bellows type actuators was reported by K. Yamamoto [3]. T. Noritsugu and H. Kobayashi also developed power
assisted wears using McKibben type pneumatic artificial muscles to increase the force for welfare workers andheavy labors[4-7]. A pneumatic soft actuator is useful for, not only power assisted system but also a rehabilitation
device. According to Japanese law, Occupational Therapist (“OT” for short) and Physical Therapist (“PT” forshort) can perform the rehabilitation to the patient for just 30 minutes per day. Some of them recommend patientsto execute voluntary rehabilitation and exercise at home. According to a medical report, a voluntary rehabilitation
might prevent bedridden state, various disease and disuse syndrome[8]. Therefore, it is necessary to develop ahome-based rehabilitation device that can safely give a passive exercise to patients. As research on such arehabilitation device, H. Taniguchi and H. Kawasaki developed a rehabilitation device for inhibiting contractureof finger joints [9-10]. In the previous study, the washable portable rehabilitation device using Extension type
Flexible Pneumatic Actuators (we call it “EFPA” for short) that can provide passive exercise was proposed andtested[11]. The attitude control of the device based on the model without position sensor was carried out. As arehabilitation device using the same EFPA, W. Tian developed the tetrahedral-shaped soft robot arm for wrist
rehabilitation was developed[12]. Usually, a pneumatic actuator was often used for wearable and rehabilitation
devices because it has advantage of compliance based on air compressibility. Especially, a pneumatic soft actuator
is more suitable because it has more advantages such as lighter weight and more compliance based on flexibility.
In the next step, we consider to develop the rehabilitation device that includes an amusement element. Typicalamusement device at home is a video game. Especially, Japanese game maker “Nintendo” produced the fitness
device with video game “Wii Fit™” in 2007. The device was also used as a rehabilitation device of outpatients
following total knee replacement[13]. It seems that the rehabilitation device with amusement is useful to carry out
the medical treatment while a patient is enjoying. In consideration of application of rehabilitation into the video
game, most of video game device only gives vibration as feedback signal to a user from a virtual space. If a device
that can safely give physical feedback movement to a patient can be realized, it is truly useful. It is also useful as1

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an amusement device. As a feedback device from a virtual space, a 3-dimantional simulator such as a flight
simulator and an automobile drive simulator had been developed as an arcade game device. However, the typical3D simulator is bulky and expensive. If a compact and safe home-based 3D simulator can be realized, the device
will be valid for both rehabilitation and amusement. In this study, we imagine a “mobile cushion” as a target
device. It is because it does not need bulky frame and can be used by just seating. In this paper, a compact 3D
simulator that can give radial inclined angle, rotational and translational motions and vibration to a seat withpatient is proposed. In detail, the development of a low height mobile robot that can rotate and translate toward
radial direction using EFPAs will be described.

PREVIOUS TETRAHEDRAL-SHAPED SOFT ROBOT ARM USING EFPAS[12]To realize a compact cushion type mobile robot, a compact soft actuator with a long stroke motion is required. In
the previous study, as a low-cost soft actuator with long stroke, the EFPA sing this type of hose was proposed andtested[11,12]. Figure 1 shows the view and schematic diagram of the EFPA. The EFPA consists of a silicone
rubber tube with inner diameter of 8 mm, outer diameter of 11 mm and covered with a bellows type nylon sleeve(The Fit Life Co. Ltd., Hose Reel). The original length of EFPA is 200 mm, and the mass is 50 g. The actuatorcan be constructed at low cost, the material cost of the EFPA is about 5 US dollars per 1 m. The EFPA can extend
about 2.5 times of its original length. The maximum pulling force of the EFPA is about 60 N[11]. The pushingforce of the original EFPA is not available because the bending stiffness is too low.

Figure 1. View and schematic diagram of EFPA[11].

Figure 2 (a) shows the appearance of the tetrahedral-shaped soft robot arm for wrist rehabilitation developed in
the previous study. In the design of the robot arm, it was intended to realize a robot arm that can bend 90 deg. andgenerate sufficient torque. The tetrahedral-shaped soft robot arm consists of three EFPAs with an original lengthof 200 mm and 22 PET sheets with a 1 mm thickness as a restraint plate. Three EFPAs are set on the three sidesof the tetrahedron except at the bottom, and they are restrained by 22 PET sheets to increase the bending stiffness.
Figure 2 (b) shows the drawings and dimensions of each reinforced PET sheet. The position of the holes for
placing the EFPA on the PET sheet are changed between 11 mm at the minimum radius and 126.5 mm at the
maximum radius from the center of each PET sheet. Each EFPA can be placed so that the EFPA can have an
elevation angle of 45 deg. with respect to the base. By restraining the three EFPAs each other with an elevation
angle of 45 deg. by many PET sheets, the axial elastic force of the rubber tube in each EFPA can be shared in thehorizontal and vertical directions. Consequently, the stiffness in both directions can be enhanced.

(a) Tetrahedral-shaped soft robot arm(b) Reinforced PET sheet

Figure 2. Appearance of the tetrahedral-shaped soft robot arm and drawing and dimensions of a PET sheet[12].

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Figure 3 shows the view of bending motion using the previous robot arm. The operating principle of the robot
arm is as follows. When one or two EFPAs are pressurized, driven EFPAs can extend and the bending motion
toward various directions along a certain sphere can be realized as shown in Figure 3. Furthermore, when all three
EFPAs are pressurized, the robot arm can extend approximately twice their length in the vertical direction. Froma preliminary experiment, the tested robot arm can bend more than 90 deg. and generate a torque of 7 Nm that canbe obtained from the experiment based on the law of virtual work.

Figure 3. View of bending motion using the previous robot arm. [12].

IMPROVED SOFT ACTUATOR AND FUNDAMENTAL CHARACTERISTICS
In this study, we consider to develop a mobile robot that uses the tetrahedral-shaped soft robot arm as a leg. In
order to apply the robot arm mentioned above to the leg, it is necessary to make the size of the actuator smaller todecrease the height of the mobile robot, and the generated torque was investigated as a fundamental characteristic.
Figure 4 (a) and (b) show the improved Tetrahedral-shaped soft Actuator (we call it “TSA” for short) and theshape of restraint PET sheets with the setting angle of EFPA of 45 and 60 deg. from the base stage, respectively.
In order to get higher stiffness and generated force, compared with the previous soft robot arm, the shorter EFPA
with the original longitudinal length of 180 mm was used for TSA. Each TSA has 10 restraint PET sheets. As
shown in Figure 4, the setting position of each EFPA are changed so that the setting angle of EFPA from the basecan keep 45 and 60 deg. Each PET sheet is installed every three pitches. The lengths of one side of an equilateraltriangle-shaped base stage using the setting angle of 45 and 60 deg. are 230 and 135 mm, respectively. The heights
of TSA using the setting angle of 45 and 60 deg. are 120 and 145 mm, respectively.

(a) EFPA setting angle 45 deg.

(b) EFPA setting angle 60 deg.

Figure 4. Tetrahedral-type soft actuator and shape of restraint PET plates.

(a) EFPA setting angle 45 deg.

(b) EFPA setting angle 60 deg.

Figure 5. Relation between bending angle and bending torque of TSA.

Figures 5 (a) and (b) show the result of bending stiffness according to applied pressure by using the TSA with
EFPA setting angle of 45 and 60 deg., respectively. In the experiment, under the condition when the base stage
was locked on a table, the horizontal pulling force was applied to the top of the tetrahedral-type soft actuator until

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The 11th JFPS International Symposium on Fluid Power 2020a certain bending angle occurs as shown in right figure on the graph. Each graph in Figures 5 shows the relation
between the bending angle and bending torque of TSA. The torque was calculated by the height of TSA and themeasured horizontal pulling force at the top of the TAS. In Figure 5, each line shows the difference of appliedpressure for three EFPAs. From Figure 5 (a) and (b), it can be seen that the bending stiffness corresponding to theslope of each line is almost same in both TSAs. It can also be confirmed that bending stiffness is increasing
according to applied pressure.
Figures 6 shows the relation between the applied pressure for two EFPAs and generated torque of both TSAs. In
Figure 6, the red and blue lines show the results using TSAs with EFPA setting angle of 45 and 60 deg.,
respectively. In the experiment, the generated torque was calculated by measured force using the experimental
setup as shown in left upper figure on graph. From Figure 6, it can be seen that the generated torque of TSA withthe setting angle of 60 deg. is superior than the case using TSA with the setting angle of 45 deg. As a result offundamental experiment using both TSAs, it can be concluded that the TSA with EFPA setting angle of 60 deg.
is suitable to apply the mobile robot from view point of almost same bending stiffness, larger generated torque.
In addition, we confirmed that TSA with the setting angle of 60 deg. can give larger bending angle anddisplacement from preliminary experiments. In addition, we also confirm that a TSA can lift up the load withmore than 130 N, that is enough lifting force for the target robot, under the condition when three EFPAs arepressurized. It means that the TSA can be used as a leg of the robot, because the TSA can bend while lifting aload when one or two EFPAs are pressurized.

Figure 6. Relation between applied pressure and generated torque.

6-LEGGED MOBILE ROBOT USING TSAS
Figure 7 shows the side and behind view of a 6-legged mobile robot using TSAs. The mobile robot consists of six
TSAs. Six TSAs are connected each other so that they can shape regular hexagon. In detail, each corner of triangleshaped base stage of TSA is connected each other by being sandwiched and screwed by two POM plates in order
to be connected surely. Adjacent TSA vertices are connected by a plastic connector. The tested robot has thelength of 410 mm, the width of 360 mm and the height of 115 mm. The planer shape of the mobile robot is theregular hexagon with a side of 210 mm. The mass of the robot is 2.85 kg including 18 valves, an embedded
controller and so on. In Figure 7, three-character Alphabet symbols indicate the location of EFPA.

Figure 7. Appearance of 6-legged mobile robot using TSAs (side and behind views)

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Figure 8 shows the schematic diagram of the control system using the tested mobile robot and motion pattern of
one leg. The control system consists of the tested robot, 18 on/off valves (Koganei Co. Ltd., G010E-1), an
embedded controller (Renesas electronics Co. Ltd., SH7125) and a serial communication unit (FTDI Co. Ltd.,FT234X). The operation of the mobile robot is as follows. First, an operator sends a character cord to the embedded
controller through PC and the serial communication unit. The embedded controller selects the sequential program
according to the received character cord. The controller also drives on/off valves according to the selectedsequential program through I/O ports. The right side figure on Figure 8 shows the moving patterns of one leg inthe robot. In the figure, the red marked circle and arrow show the pressurized EFPA and bending direction,
respectively. By changing the pressurized EFPA, the leg can bend in six radial directions every 60 deg. By
pressurizing all EFPAs, it can also extend about double toward the longitudinal direction for lifting the robot.

Figure 8. Schematic diagram of control system on the tested mobile robot and motion patterns of one leg.

DRIVING TEST OF THE MOBILE ROBOT USING TSAS
In order to realize the propulsion of the mobile robot, it is necessary to adjust the driving pattern of each TSA.
The method of propulsion is as follows. While lifting the whole robot body by extending three TSAs (Supportlegs), other TSAs bend toward the propulsion direction as a free leg. After that, while lowering the robot body bycontracting the extended TSAs, the other TSAs are bending toward the opposite direction. By this method, therobot can move toward six directions based on the motion patterns of the leg as shown in Figure 8. In addition, as
three legs works for lifting up the load, the robot can convey a mass with about 40 kg theoretically.

(a) Time chart of driving pattern.
(b) Transient view of movement.
Figure 9. Time chart of driving pattern of each EFPA and transient view of movement of the robot for moving forward.

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Figure 9 (a) shows the time chart of sequential driving pattern of each EFPA and transient view of movement of
the robot for moving forward. In the time chart, blue bars show the driving pattern for lifting the robot body. Redbars show the bending motion. In this sequence, the combination of the motion with lifting and swing motion fora free leg and the motion with lowering and swing motion for kicking are repeating two times. In addition, eachmotion is carried out for the time interval T. In the case of moving forward, EFPAs related to moving forward are
five in three TSAs, that is FFB, BRL, BRB, BLB and BLR. Figure 9 (b) shows the transient view of movementof the robot in the case of moving forward. In the experiment, switching timing of each motion T of 360 ms wasused. From Figure 9 (b), it can be seen that the robot can move forward smoothly.
Figure 10 show the driven EFPAs for swing motion of legs in moving toward six directions, circled number shows
the moving direction of the robot. Symbols ● show the driven EFPA. From Figure 10, it can be seen that six
different EFPAs are driven every moving direction.

Figure 10. Driven EFPAs for swing motion of legs in moving toward six directions.

In order to obtain the optimal switching time T, the driving test of the robot for moving forward with various
switching timing was carried out. In the experiment, the averaged moving distance for one driving pattern and theaveraged velocity of the robot, when the moving pattern as shown in Figure 9 was repeated three times, were
investigated. As a preliminary experiment, the moving distance and the velocity of the robot were investigatedwhen the switching time T was changed from 50 to 500 ms every 50 ms. In addition, to obtain the optimal
switching time, both distance and velocity were also investigated more detail between 300 and 400 ms every 10
ms. Figures 11 (a) and (b) show the relation between switching timing T and the moving distance for one driving
pattern and velocity of the mobile robot in case when the timing T is from 50 to 500 ms every 50 ms and thetiming T is from 300 to 400 ms every 10 ms, respectively. In both figures, the red and blue lines show the moving
distance and velocity of the robot, respectively. It can be seen that the maximum moving distance can be obtainedby using the switching timing of 370 ms. About a velocity, it can be seen that the velocity becomes peak aroundboth switching times of 310 and 360 ms. From the view point of longer stroke and higher velocity, it can beconcluded that the switching timing T of 360 ms is more suitable to the moving forward.

(a) T is from 50 to 500 ms every 50 ms.

(b) T is from 300 to 400 ms every 10ms.

Figure 11. Relation between switching timing T and moving distance for a driving pattern and velocity of the mobile robot.

Figures 12 (a) and (b) show time charts of driving pattern of each EFPA when the robot rotates in clock-wise andcounter clock-wise, respectively. The switching timing T1 and T2 were adjusted by trail and errors. In theexperiment, T1 of 400 ms and T2 of 350 ms were used, respectively. As a same manner of moving forward, in therotational operation, three TSAs work to rotate the robot. Figure 13 shows the transient view of rotational motion

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The 11th JFPS International Symposium on Fluid Power 2020of the mobile robot when the robot rotates in counter clock-wise. In the experiment, the rotational speed of 15 deg.
can be obtained. As a result, it can be concluded that the mobile robot for a home-based compact 3D simulator
that can give rotational and translational motions can be realized. In addition, we also confirm that the robot canchange the radial inclined angle of the robot and give a vibration.

(a) Clock-wise.

(b) Counter clock-wise.

Figure 12. Time chart of driving pattern of each EFPA for rotational motion in clock-wise and counter clock-wise

Figure 13. Transient view of movement of the robot for rotational motion.

CONCLUSIONSThis study that aims to develop the mobile robot that can be used for core training and rehabilitation can besummarized as follows.
In order to investigate the suitable Tetrahedral-Shaped soft Actuator (TSA) for the desired mobile robot, two typesof TSAs with EFPA setting angle of 45 and 60 deg. were proposed and tested. The bending stiffness and generated
torque using both TSAs were investigated. As a result, it can be concluded that the TSA with setting angle of 60
deg. is suitable to the mobile robot because of larger stroke and higher torque. The mobile robot that consists of
six TSAs was proposed and tested. The control system that consists of the tested robot and 18 on/off valves was
also proposed and tested. The sequential driving pattern for moving forward and rotating in clock-wise and counterclock-wise was investigated. As a result, it can be confirmed that the tested robot can move toward six directions
and rotate toward both clock-wise and counter clockwise. In addition, the optimal switching timing for moving
forward was investigated. As a result, it was found that the switching timing of 360 ms is suitable to a moving
forward. In the tested robot, the maxim velocity of 80 mm/s and the rotational speed of 15 deg/s can be obtained.

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ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Numbers 16K06202 and 19K04265.

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2. NAGATA, Y., Soft Actuators, Forefront of Development. NTS Ltd., 2004, pp. 291–335.
3. ISHII, M., YAMAMOTO, K. and HYODO, K., Stand-Alone Wearable Power Assist Suit (Development and Avail ability).
Journal of the Japan Society of Mechanical Engineers, 2006, 72-715, pp. 175–182.
4. NORITSUGU, T., Pneumatic Soft Actuator for Human Assist Technology. In: Proceedings of the 6th JFPS International
Symposium on Fluid Power, Tsukuba, Japan, 2005, pp. 11–20.
5. SASAKI, D., NORITSUGU, T., YAMAMOTO, H., and TAKAIWA, M., Development of Power Assist Glove using
Pneumatic Artificial Rubber Muscle. Journal of the Robotics Society of Japan, 2006, 24-5, pp. 640-646.
6. NORITSUGU, T., TAKAIWA, M. and SASAKI, D., Development of Power Assist Wear Using Pneumatic Rubber
Artificial Muscles. Journal of Robotics and Mechatronics, 2008, 21-5, pp. 27–31.
7. KOBAYASHI, H., SHIBAN, T., and ISHIDA, Y., Realization of all 7 motions for the upper limb by a muscle suit. Journalof Robotics and Mechatronics, 2004, 16-5, pp. 504–512.
8. Council of Bureau of Social Welfare and Public Health Tokyo Metropolitan Government from, https://www.
fukushihoken.metro.tokyo.lg.jp/iryo/sonota/riha_iryo/kyougi01/rehabiri24.files/siryou242.pdf; February 2013 (inJapanese).
9. TANIGUCHI, H., WAKIMOTO, S., SUZUMORI, K., Development of a hand rehabilitation system to prevent contracture
for finger joints based on the therapy of occupational therapists (Massage a hand and range of motion exercises using
pneumatic soft actuators). Journal of the Japan Society of Mechanical Engineers, 2014, 80-820, TRANS0348.
10. KAWASAKI, H., ITO, S., ISHIGURA, Y., NISHIMOTO, Y., AOKI, T., ABE, T., SAKAEDA, H., OJIKA, T., MOURI,
T., UEKI, S., Hand Rehabilitation Assist Robot Designed with Assistance for Thumb Opposability. Journal of the Japan
Society of Mechanical Engineers, 2008, 74-748, pp. 3019-3027.
11. SHIMOOKA, S., HANE, Y., AKAGI, T., DOHTA, S., KOBAYASHI, W., SHINOHARA, T., and MATSUI, Y.,“Development and Attitude Control of Washable Portable Rehabilitation Device for Wrist without Position Sensor”, JFPSInternational Journal of Fluid Power System, Vol.13, No.3, 2020, pp. 25-34.
12. TIAN, W., JHAN, C., INOKUMA, M., AKAGI, T., DOHTA, S., and SHIMOOKA, S., ”Development of a TetrahedralShaped Soft Robot Arm as a Wrist Rehabilitation Device Using Extension Type Flexible Pneumatic Actuators”, Journalof Robotics and Mechatronics, Vol.32, No.5, 2020, pp. 931-938.
13. FUNG, V., HO, A., SHAFFER, J., CHUNG, E., GOMEZ, M., “Use of Nintendo Wii Fit™ in the rehabilitation of
outpatients following total knee replacement: a preliminary randomised controlled trial”, Physiotherapy. 2012, 98-3, pp.
183-8.

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Development of the Transfer System for Bedridden Elderly andDisabled People using Pneumatic Actuators
Feifei CHO* , Keisei KATO*, Ryota ENDO* ,Takumi KOBAYASHI** and Tetsuya AKAGI***

Dept. Mechanical Systems, National Institute of Technology Tsuyama College624-1,Numa, Tsuyama City, Okayama, 708-8509, Japan(E-mail: cho@tsuyama-ct.ac.jp)**
Dept. Intelligent Mechanical Engineering, Okayama University of Science

Abstract. In recent years, as the birthrate declines and the population ages in developed countries, the demand for
caregivers has been increasing. However, nursing care often involves a physical burden on the caregiver, especially intransfer assistance, which places a burden on the entire body. To solve this problem, this study aims to develop a novel
support system such as automatic pneumatic transfer sheets, which can help caregivers to move bedridden elderly
people bed-to-bed and bed-to-fully reclining wheelchair. In addition, by using a pneumatic actuator in the roller partand controlling the air pressure in time, this device can not only change the installation surface between the sheet andthe back, but also can be expected to prevent bedsores. So far, we have built a prototype, which is half the size of theactual transfer device by using a low-cost servo valve using buckled tubes.
Keywords: Air pressure, Flexible artificial muscle, Transfer system, Servo valve using buckled tubes

INTRODUCTION
In recent years, the number of bedridden elderly people in the world has been increasing remarkably due tothe aging of society, and the number of disabled people who fall into severe physical disabilities due to various
diseases is not small. In addition, caregivers at home are aging in Japan, as exemplified by the elderly caresystem[1]. In caring for these elderly and disabled people, transferring them from bed-to-bed and bed-to-fully
reclining wheelchair is a heavy physically demanding task for caregivers. Also when caring for bedridden persons,caregivers have to change their posture every two hours because it is necessary to prevent bedsores. In some cases,
the heavy burden on caregivers’ back and arms might make them uncomfortable, and then increases their mental
burden[2].
The above physical and mental burdens are major factors that lead to violence and harassment in nursing
homes and home care. These problems return to society as the burden on long-term care insurance premiums, andthus have a significant impact on economy. In order to solve the above problems, the transfer devices and power
assist devices are in high demand. However, most of developed devices such as floor traveling lifts, which requirelarge amount of power to lift the care receiver, are expensive and moreover they require the time to wear belts
and the labor to hold the care receiver’s body. There are also practical low-cost sliding seats sed in homes andnarrow facilities, but they require two or more people to support the care receivers, and require a lot of energy toslide them. The same as the floor traveling lifts, caregivers have to take lots of time to prepare the seat under thecare receivers. Those answered by more than 50% of the respondents were mental burden 74.3%, temporal
restriction 62.5%, and physical burden 53.3%[3]. Figure 1 shows the difficulties in providing care at home.

Figure 1. Difficulties in providing care at home[3]Therefore, in this study, we aim to develop the low-cost flexible transfer and repositioning system for
bedridden elderly and disabled people, which can be used for the bed sheet ordinarily. This system also can

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generate a large transfer force using soft actuator. They consists of a microcomputer, servo valves using buckled
tubes[4-5] and soft pneumatic actuators.

OVERALL IMAGE OF THE DEVICE
In this study, we aimed at the improvements of the floor traveling lifts and the slide seat, which were describedin the previous chapter, and designed the transfer system for bedridden elderly and disabled people using
pneumatic actuators. Figure 2 shows the overall image of the device. The design concept is as follows:(1) One caregiver can support a care receiver by herself/himself.
(2) No more physical burden on the caregiver.
(3) Easy to prepare, and it will not take long time.
(4) No need to change the bedridden elderly people posture every two hours.
An explanation of the concept is given below. In most cases of floor traveling lifts and sliding seats, one
caregiver has to transfer the bedridden elderly people and the other caregiver has to support the care receiver tomake sure that the movement will be safe. Therefore, we proposed the powered sliding seat that can transfer
automatically. Then, by using the McKiben actuator for the roller part, it will be realized. Because, when theactuator is pressurized, the actuator becomes stiff such as the roller. Then, when the actuator is exhausted, theactuator returns to the flexible state. It does neither take much time to prepare, nor need a lot of space. The
principle of operation of this system is as follows. We put the electric motors at both ends of each actuator, then,by rotating the motors, the actuators also rotate, and the sheet moves simultaneously. The bedridden elderly people
will be transferred in the manner of the roller conveyor. In order to maintain the flexibility, we used soft materials
for a sheet. The operating principle is as follows: when the McKibben actuator is pressurized, the actuator isexpanded in radial direction, and it changes from the flexible state to the slightly stiff roller. By controlling theactuators in time, the contact surface between the human back and the bed sheet can be changed, thereby reducing
the steaminess of the back and suppressing bedsores. This system can be expected to not only reduce the frequencyof repositioning but also reduce the burden on caregivers. Figure 3 shows the construction of the connector partof the actuator. Because of the possibility of contact with the care receiver body, most of ordinary connectors are
made of metal, which is inferior in terms of human affinity. The muscle is also contracted about 25% of theoriginal length in axial direction. The average height of elderly people over 70 years old is about 1.6m, therefore,
when transferring the caregivers, their knees should be bent.

Figure 2. Overall image of the device

Figure 3. Construction of the connector part of the actuator
First, the internal pressure required for one actuator is determined. The number of actuators supporting the backis obtained from Equation (1), the external pressure applied to one actuator is obtained from Equation (2).
Determine the internal pressure and external pressure of the hose is obtained from relational Equation (3). The

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required torque is obtained by Equation (4). This value does not include the internal pressure required to expand
the silicon tube. As shown inTable 1, the parameters used in the calculations are h=0.411m, W=53.4kg, d=0.03m,
r=0.03m, and l=1.4m. Where, W and h are the average weight and shoulder width of the care receivers over theage of 70, respectively. The value of W was calculated by back calculating the value of BMI(Body Mass Index),
which is the maximum possible external pressure. We used the maximum value 30 of BMI and the average heightof the care receivers over the age of 70. As a result, W =77.8kg. Table 1 shows the symbols and parameters usedin the calculation of theoretical values.
MAXℎ𝑛 = (𝑑+𝑟)
𝑃𝑒 =
𝑃𝑖 =

𝑊𝑔
𝑛𝑙𝑑
2𝑃𝑒𝑡
𝑑

1

𝑇 = 2𝑛 𝑑(𝑊𝑔)DiameterThickness
LengthSpacing

(1)
(2)
(3)

(4)

Table 1. Symbols and parameters be used in the calculation of theoretical values[6-7].
The care receivers(Over 70 years old)
Actuator(Pressurized)
Numbernd
0.03m
Average weightW
53.4kgInternal pressurePit0.003mExternal pressure
Pe
l
1.4m
Average shoulderh0.411mwidth
TorqueTr
0.03m

As a result, assuming the average shoulder width of bedridden elderly people who require nursing care over theage of 70, we used eleven McKiben actuators to support the human body, the external pressure applied to one
actuator will be about 2MPa, and the required internal pressure was about 0.4MPa. The pressure resistance of thesilicon tube is about 0.4 MPa with thickness of 0.003m, and this value is sufficient to withstand the pressure. Then,the torque required for one actuator was calculated to be 1.31N・m. This value does not take into account thefriction of the sheets and gears. Figure 4 shows the device of the transfer system.

Figure 4. The device of the transfer system

CONTROL SYSTEM
In this system, it is necessary to apply the air to the pneumatic actuators during transferring, rotate the motors,
and control pressurizing the actuators in time because it can reduce back steaminess and suppress bedsores forbedridden elderly people. The SH7125 microcontroller is used for this system. Figure 5 shows the schematic
diagram of the control system. The control system consists of a tested actuator, a control valve, a pressure sensor
and an embedded controller. In consideration of controlling multiple pneumatic valves, we expanded the numberof pins capable of PWM output from 4 to 8 that will be allowed for more output. In addition, pneumatic valves
are required to operate the pneumatic actuators. However, the pneumatic control valve is the most expensive
control device in the fluidic drive system, and it is the big obstacle to make the support device usable at home.
Therefore, we used the servo valve using buckled tubes as shown in Figure 6. The valve consists of two buckled
soft polyurethane tubes (SMC, TUS0425), two one-touch connectors (Koganei, US4M), one Y-shaped one-touchconnecters (Koganei, UY4M), a small-sized RC servo motor (Asakusa Giken, ASV-15: Max. torque of 1.6kgfcm, Max. speed of 480deg./s, band width frequency of 3 Hz in experiment) and an acrylic rotational disk withconnecter holders. The mass of the valve is 40g. The valve has a length of 45 mm, a width of 79 mm and heightof 53 mm. Finally, we checked the response of the inner pressure of the actuator by using PI control scheme and3

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the servo valve using buckled tubes. Figure 7 shows the transient response of the pressure. From the figure, it canbe seen that the response is close to the stepwise target pressure of 400KPa.

Figure 5. Schematic diagram of the control system

Figure 6. Servo valve using buckled tubes

Figure 7. The result of PI control

CONCLUSIONSIn this study, we aimed to develop the transfer system for bedridden elderly and disabled people using
pneumatic actuators and we also constructed the control system. In the prototype, we have tried the method oftransmitting power from the motor using the bearing to rotate the actuators. We also used SH7125 microcomputer
to control the system. In the future work, we plan to improve the device with the new transport method. Finally,
we have got confirmation about the movement of this transfer system by using the low-cost servo valve using
buckled tubes.

REFERENCES
1. Cabinet Office, Government of Japan: White Paper on the Aging Society, 2009, Section 1: Status of Aging,https://seniorguide.jp/article/1207592.html
2. Fujimoto T., Higashi Y., Takahashi S., Yasuda T., Tamura T., Current status and future of nursing care equipment infacilities, J-STAGE, 2000,14-2, 36-45.
3. LIFULL: The Burden of Struggling with Residential Care, https://kaigo.homes.co.jp/manual/voice/kentou/kurou/
4. Kbayashi H., Akagi T., Dohta S., Kobayashi W., Matsui Y., Shimooka S., Shinohara T, and Aliff M., "Development ofPortable Rehabilitation Device Driven by Low-Cost Servo Valve Using Tap Water" International Journal of MechanicalEngineering and Robotics Research, Vol. 9, No. 3, pp. 353-359, March 2020. DOI: 10.18178/ijmerr.9.3.353-359
5. Shimooka S.,Akagi T., Dohta S., Matsui Y., Shinohara T, Hane Y.,and Aliff M., "Development of Reinforced Extension
Type Flexible Pneumatic Actuator with Circumferential Restraints and Its Application for Rehabilitation Device"
International Journal of Automotive and Mechanical Engineering. 17(3), 8116-8127. July-September 2020.DOI:
https://doi.org/10.15282/ijame.17.3.2020.05.0609.
6. Ministry of Economy, Trade and Industry: Average of key dimensions by age group,http://warp.da.ndl.go.jp/info:ndljp/pid/
286890/www.meti.go.jp/press/20071001007/004_bessi.pdf
7. Ministry of Health, Labour and Welfare: Results of Physical Condition Survey, https://www.mhlw.go.jp/content/0004517
60.pdf4

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Development of Pneumatic Drive Pipe Inspection Robotusing Radial Bending Type Soft Actuator
Takashi SHINOHARA*, Hikaru FURUYA**, Tetsuya AKAGI **, Shujiro DOHTA **,Takumi KOBAYASHI** and So SHIMOOKA******

Science Dream lab., Okayama University of Science
Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science
1-1, Ridai-cho, Kita-ku, Okayama, 700-0005 Japan(E-mail: shinohara@pub.ous.ac.jp)***
Graduate School of Natural Science and Technology, Faculty of Engineering, Okayama University
1-1-1 Tsushima-naka, Kita-ku, Okayama, 700-8530 Japan

Abstract. In this paper, for the purpose of developing a pipe inspection robot that can catch pipes whose diameterhas changed due to rust generated inside, the pipe holding mechanism that can flexibly catch various size of pipes
was prototyped. A novel pipe holding mechanism using three extension type flexible pneumatic actuators (EFPAs)
that holds the pipe while bending radially was proposed and tested. Also, the holding characteristics for various
pipes of the tested mechanism is investigated. Furthermore, the inchworm-type robot using the tested holding
mechanism was also proposed and tested. As a result, it could be confirmed that the robot could travel on the straightpipes and complicated pipes that consists of different pipes. In addition, a novel propulsion mechanism using only
two holding mechanisms was also proposed and tested. It was also confirmed that a speed of 4.5 mm/s was obtained
and that it could travel in a curved pipe.
Keywords: Pneumatic Drive Pipe Inspection Robot, Extension type flexible pneumatic actuator, Radial bending
type soft actuator, Radial curved type pipe holding mechanism.

INTRODUCTION
In 2014, about 12.1% of the total length of water supply pipeline (660,000 km) in Japan has already passed theendurance period of 40 years [1]. As only 0.76% of the total pipe is changed every year, it will need more than
130 years to renewal the whole pipeline. In addition, water pipelines are complicated and consist of numerous
corners and joints. An excavation work to inspect the entire pipes is expensive and time consuming. Therefore,various inspection methods such as a fiber scope, a dowsing and a wheel type inspection robot have been
proposed and developed. However, a manipulated fiber scope cannot travel with long distance because oflimitation of pushing force. Consequently, various companies and researchers have developed various selfpropelled pipe inspection robots [2]-[8]. By using these robots, the cost of pipe inspection can also be reduced.
Hence, inspection robots are required to have high mobility in the pipe. In ideal, the shape of the robot bodychanges naturally in order to consume less energy and shorten the travel time. Therefore, several researchersdeveloped inchworm type pipe inspection robots using pneumatic soft actuators such as pneumatic artificial
muscles and flexible cylinders [5]-[8]. In addition, a fluidic actuator does not use electricity as the primaryenergy, and because there is no risk of electric leakage or short circuits, they can be safely used in damp areasuch as in water supply pipe [5]-[12]. Based on this concept, we have proposed and tested an inchworm typepipe inspection robot that consists of a sliding propulsion mechanism and a holding mechanism using extension
type flexible pneumatic actuators (it is called “EFPAs” for short) in our previous study [13]-[14]. The robot wasable to move smoothly in a complex pipe with the inner diameter of 100 mm [14].
However, the tested pipe holding mechanism can only hold a pipe with a certain diameter [14]. In the next step,for developing a pipe inspection robot that can travel in various sized pipes by using only one type pipe holding
mechanism, a novel pipe holding mechanism that can flexibly catch pipes with various inner diameters is
proposed in this paper. In detail, a radial curved type pipe holding mechanism using three EFPAs that can catchthe pipe while bending radially is proposed and tested. Also, the holding characteristics for each pipe (150A,
100A, 75A standards) of the prototyped pipe holding mechanism is investigated. In addition, the inchworm-type
pipe inspection robot using the tested holding mechanisms is proposed and tested. Furthermore, we propose anovel propulsion method using only two radial curved type holding mechanisms arranged back-to-back, andinvestigate the optimal driving timing of the front and rear holding mechanisms for propulsion by experiments.

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PREVIOUS INCHWORM TYPE PIPE INSPECTION ROBOTIn the previous study, the inchworm type pipe inspection robot using EFPAs was developed [13-15]. Figure 1
shows the construction of the extension type flexible pneumatic actuator (EFPA) [16-18]. The EFPA consists ofa silicone rubber tube with an inner diameter of 8mm and outer diameter of 10mm covered with a bellowsshaped fabric sleeve made of nylon strings and two end connectors made of acrylic with a supply port. The
operating principle of the EFPA is as follows. When the supply pressure is applied to the chamber, the innerrubber tube expands toward radial and axial directions. As the bellows-shaped sleeve prevents to expand toward
radial direction, the EFPA extends longitudinally. The EFPA can extend more than about 2.5 times from its
original length.

Figure 1. The construction of the extension type flexible pneumatic actuator (EFPA).

Figure 2 shows the construction of sliding/bending mechanism using three EFPAs developed in the previous
study [14]. In the mechanism, three EFPAs are set in parallel between both disks in every 120 degrees and havea radius of 12 mm from the central axis of the robot. The operating principle is as follows. The expansion andcontraction operation are realized as follows; the mechanism is extended when three EFPAs are pressurized at
the same time. And the mechanism is returns to the original shape due to the elasticity of them when three
EFPAs are exhausted simultaneously. In addition, the bending motion is applied by pressurizing one or two
EFPAs in the sliding/bending mechanism.

Figure 2. The construction of Sliding/bending mechanism using three EFPAs.

Pipes with smaller inner diameter of 75 and 50 mm that are commonly used as main pipe diameter were selected
as a tested pipe in this experiment to a tire wheel type pipe holding mechanisms. In the previous study, in orderto get a larger contact area with pipe without a larger volume change, a tire wheel type pipe holding mechanism
for 75A standard pipe as shown in Figure 3 (a) is proposed and tested. The mechanism consists of a tire wheel
type cylindrical rubber tube holder made by laminated acrylic plates, the rubber tube with outer diameter of 60
mm and two disk shaped holding plates and clamping rings for rubber tube. In the mechanism, the rubber tube is
set over the wheel type holder. Both ends of the tube are clamped by the tube holder and disk-shaped holding
plates through two clamping rings. By this method, the volume change of inner chamber surrounded by rubber
tube for input pressure is limited toward only radial direction of the mechanism. The mass of the tested
mechanism is 68.8 g.

(a) Construction(b) Pipe holding characteristics
Figure 3. The construction and pipe holding characteristics of the tire wheel type pipe holding mechanisms.

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Figure 3 (b) shows the relation between supplied pressure and pipe holding force of tested mechanism for pipe
diameter of 75A. In the experiment, the pipe holding force is measured as a pulling force when the slip has
occurred under the condition when the mechanism in the pipe is pulled. In the figure, symbol shows the averagevalues of holding force of 20 times measurements. The vertical bar shows the measured data distribution. It canbe seen that the tested mechanisms can generate pipe holding force of about 160 N. This is enough to apply topipeline inspection robot using EFPAs, because the maximum generated pulling force of the sliding/bending
mechanism in the robot is about 60 N.
Figure 4 shows the construction of the improved pipe inspection robots using the improved holding mechanism.
Two holding mechanisms are set to both ends of the sliding/bending mechanism. In both sides of the holding
mechanism, there are plastic arc-shaped guide to prevent stacking at a corner as shown in Figure 4.

Figure 4. The construction of the tire wheel type pipe holding mechanisms.

The operating principle of the robot is as follows. First, the pipe holding mechanism at the end of the robot
expands so that it can hold the pipe. In the condition, the sliding mechanism is extended by pressurizing three
EFPAs. When the sliding mechanism extends up to a maximum length, the top pipe holding mechanism
expands to hold the pipe. After that, the end pipe holding mechanism and the sliding mechanism contracts by
exhausting input pressure, and end holding mechanism moves forward. By repeating these operations, the robot
can move forward as an inchworm. The backward motion can also be realized by applying the opposite
operation mentioned above. To steer the robot toward desired direction in a pipe joint, after bending motion isapplied by pressurizing one or two EFPAs in the sliding/bending mechanism, it can be realized by all three
EFPAs are pressurized. By this method, the robot can select six radial directions by pressurized one or two
EFPAs in the sliding/bending mechanism.
Figure 5 shows the schematic diagram of control system to drive the tested pipe inspection robot. The system
consists of the tested robot, five on/off control valves and an embedded controller. Three valves are used for thesliding mechanism, and they are operated through transistor and I/O port in the embedded controller. Twovalves are used for pipe holding mechanisms and operated through transistor and PWM control port in thecontroller. In order to use the same supplied pressure of 500 kPa for both sliding and holding mechanisms, theholding mechanisms are driven by lower duty ratio signal of 50 % so as to reduce operating pressure of themechanism. The control procedure is as follows. The operator sends digital code through a serial
communication cable using a personal computer. The embedded controller selects to drive suitable valve
according to the sequential program that is preinstalled into the controller. The timing of valve operation isadjusted by using a timer function in the embedded controller.

Figure 5. Schematic diagram of control system to drive the mechanism.

Figure 6 shows the results of driving test when the tested inspection robots pass through a complex pipe. In theexperiment, three straight pipes with a length of 1 m and a corner and a tee joint were used. In addition, thevalve unit was located on outside of pipes. Between the robot and valve unit, there is parallel arranged five small
pipes with outer diameter of 4 mm and inner diameter of 2.5 mm. From Figure 6, it can be seen that the robot

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can pass through smoothly the complex pipe with a total length of 3 m within 4 minutes. Interestingly, throughobservation, robots are not stuck in corner and tee joints. This is mainly because of the improved higher holding
force.

Figure 6. The results of driving test when the tested robots pass through a complex pipe with 75A pipe [14].

RADIAL BENDING TYPE HOLDING MECHANISM
The robot using the tire wheel type pipe holding mechanisms could travel in complex pipes with inner diametersof 75 mm. However, it could only travel in pipes with a certain diameter. The tested holding mechanism can notuse for pipes with different diameters such as from 100 mm to 75 mm or to 150 mm. When the holding
mechanism catches larger pipe with more than 75 mm, the mechanism doesn’t work and occurs explosion. In
this study, we aim to develop the pipe inspection robot that can travel in different diameter pipes of 150A, 100Aand 75A (inner diameters of 150, 100 and 75mm), they are typical water supply pipes (excluding householdbranch pipes). We also aim to develop a holding mechanism that can hold various sized pipes or multiple pipes.
The design concept of the holding mechanism for different diameter pipes is as follows. In order to hold several
different diameter pipes by using only one holding mechanism, the mechanism needs to increase the radial
expansion. Although balloon type mechanism that expands seems to be suitable, it consumes a large amount of
air. Also, if the expansion due to elastic deformation of material is applied in it, it needs to worry about bursting.
Therefore, we apply the method that the length of the material required to change in the radial direction issecured in advance in the direction of the pipeline axis with a margin. In detail, we consider a mechanism that it
can change in the radial direction like a flower. The mechanism can hold the inside pipe wall when the actuator
expands radially from the initial condition of no pressure. In the initial condition, the actuator extends in theaxial direction of the pipe. In addition, this holding mechanism needs to have some flexibility for traveling incurved pipes. Figure 7 shows the construction of a radial bending type pipe holding mechanism prototyped
based on the design concept. In the holding mechanism, three EFPAs with a natural length of 75 mm are placedin parallel at a radius of 11 mm from the center every 120 deg. In each EFPA, outer side of EFPA is restrainedby connecting both ends through plastic beam like a cantilever beam.

Figure 7. The structure of the radial bending type holding mechanism.

Figure 8 shows the attachment parts for the base that fastens the three EFPAs and the ring-shaped restraint plateat the bellows of each EFPA. The base plate mounted on three EFPAs has three holes with an inner diameter of
8 mm that are located on a radius of 11 mm from the center every 120 deg. Four small holes (for M3 screw) are
set on a radius of 7 mm from the center of each large hole so that each EFPA can be easily removed from themechanism. The ring-shaped restraint has a groove of 1 mm in length and 11 mm in width as shown in Figure 8
A POM restraint plate (thickness of 1 mm, width of 10 mm) is installed into the groove of ring-shaped restraint

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plates on the outer peripheral side of each EFPA so that the outer side longitudinal length does not change. This
ring-shaped restraint plates are installed every four bellows so that the bending motion can occur when EFPAs
extend. In addition, each end of EFPA is covered with a silicone rubber tube (with the inner diameter of 15 mm,
outer diameter of 18 mm and length of 20 mm) to prevent a slipping. The mass of the holding mechanism is 110
g.

Figure 8. Mechanical drawing of the base plate (Left side) and ring-shaped restraint plate (Right side).

Figure 9 shows the appearance of the radial curved holding mechanism when supply pressure of 500 kPa isapplied. It can be seen that the restraint plate can change the extension of each EFPA into a bending motion, andthe three EFPAs can be bent radially. The maximum diameter for the applied pressure of 150 kPa is 200 mm.
When the applied pressure becomes higher that is more than 200 kPa, each EFPA warps further if the end of
EFPA does not reach the inner wall, and the outer diameter of the mechanism become smaller. Therefore, thediameter of the holding mechanism is maximized when the applied pressure is 150 kPa. If it is used in a pipeline
with diameter of 200 mm or less, it is not necessary to adjust the pressure.

Figure 9. The appearance of the radial curved holding mechanism when 500 kPa is applied.

The relations between the applied pressure and the holding force using various sized holding mechanisms areshown in Figure 10. Figure 10 (a) shows an experimental setup using prototype holding mechanisms. Figures 10(b), (c) and (d) show results using the restraint plate with length of 40 mm (thickness of 1 mm, width of 10 mm),
75 mm (thickness of 1 mm, width of 10 mm) and 100 mm (thickness of 2 mm, width of 10 mm), respectively. In
each graph, symbols ●, ▲, and ♦ are the average values of the holding force measured 20 times in the case of
the pipe of standards 75A (inner diameter of 79 mm), 100A (inner diameter of 105 mm), and 150A (innerdiameter of 155 mm), respectively. In the experiment, the holding mechanism inserted into the pipe, and the
EFPAs was pressurized. Also, it was pulled through a force gauge (Nidec-Shimpo Corporation; FGPX-100) thatis connected in series with the holding mechanism as shown in Figure 10 (a). And the force at the time when theslip occurs is defined as a holding force. Based on 20 times measurements, the average value and distribution ofholding force were calculated. From Figure 10, It can be seen that the holding mechanism with the length of 40
mm cannot hold the pipe of 150 A, and the mechanism with length of 75 mm or more can hold all three types ofpipes. Also, the holding force using the mechanism with length of 40 mm is larger than the case using others.
On the other hands, the holding mechanism with a length of 75 mm has a good holding force that can hold three
types of pipes. Therefore, it seems that a holding mechanism with a length of 75 mm is suitable for all three
types of pipes. For 100A and 75A standard pipe, the holding mechanism with length of 40mm is suitable,
because it has the larger holding force and compact body for curved pipe. In addition, it can be seen that all
types of tested holding mechanisms have enough holding force to support the mass of the robot (less than 1 kg),except for the combination that holds the pipe 150A with the holding mechanism of 40 mm.

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The 11th JFPS International Symposium on Fluid Power 2020(a) Experimental setup for holding force.

(b) Result using the mechanism with length of 40 mm.

(c) Result using the mechanism with length of 75 mm.

(d) Result using the mechanism with length of 100 mm.

Figure 10. Experimental setup for holding force and relation between input pressure and holding force using

various pipe holding mechanisms.
Figures 11 (a), (b) and (c) show the appearance of a pipe inspection robot using the holding mechanism withlength of 75 mm, the construction of an omni wheel mechanism in the robot, and the appearance of a sharedpipeline plate for the pipe holding mechanism, respectively. The tested robot consists of a propulsion
mechanism using the extension / bending mechanism and two radial curved type pipe holding mechanisms withlength of 75 mm. The end of both holding mechanisms connected to both ends of the propulsion mechanism.
Also, the omni wheel mechanism to reduce the friction and prevent stacking on the curved pipe is installed
between the pipe holding mechanism and the propulsion mechanism. Since the holding mechanism needs three
EFPAs to bend radially at the same time, it is necessary to share a supply pipeline into three outlets. The sharedpiping plate as shown in Figure 11(c) has branches for three outlets for EFPAs from a supply port. And it can becompactly integrated. The pressurized air to the front pipe holding mechanism is supplied from the rear side byusing a coil type tube (inner diameter of 2.5 mm, outer diameter of 4 mm). The robot has a length of 520 mm, anouter diameter of 73 mm, and the total mass of the robot is about 1 kg.

(a) The appearance of the pipe inspection robot using the pipe holding mechanism with length of 45 mm.

(b) Construction of omni wheel mechanism.
(c) The construction of shared pipeline plate.
Figure 11. The structure and components of the pipe inspection robot.

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The control system of the tested robot is the same as the previous one as shown in Figure 5. Figures 12 (a) and(b) show the time chart of each valve for the robot moving forward and moving forward with bending,respectively. In Figure 12, V1 means the valve to drive the front holding mechanism. V2, V3, and V4 mean
valves to drive three EFPAs in the sliding / bending mechanism. V5 means the valve for the rear holding
mechanism. By repeating the operation of moving forward that one cycle requires 11 s as shown in Figure 12 (a),the robot can move forward continuously. The moving backward can be realized by reversing the operation of
the front and rear holding mechanism. In the propulsion in the curved pipe, that is the operation for moving
forward while bending as shown in Figure 12 (b), one or two of the three EFPAs are pressurized in order tocreate a bending motion in a certain direction before the extension operation of the sliding / bending mechanism.
After that, the three EFPAs are pressurized to extend. The timing of holding is almost the same as the case ofmoving forward, and one cycle for traveling curved pipe is 11.25 s.

(a) For moving forward (straight pipe).

(b) For moving forward while bending (carved pipe).

Figure 12. Time chart of each valve of the robot for moving in straight and curved pipes.

Figures 13 (a) and (b) show a schematic diagram of a pipeline in which different straight pipes are connected inseries and a transient view of movement of the robot in traveling test of the pipe inspection robot, respectively.
From Figures 13 (b), it can be seen that the tested robot can travel smoothly even if the pipe diameter changes
under the condition when the pipeline is set horizontally. It was found that the total travelling time of thepipeline of about 3 m was 210 s, and the average speed of the robot was 14.3 mm/s.

(a) A schematic diagram of a pipeline with different straight pipes.

(b) Transient view of movement of the robot.
Figure 13. Transient view of movement of the robot in traveling test of straight pipe.

In addition, Figure 14 (a), (b) and (c) show the schematic diagram of the complex pipeline, the transient view ofthe movement of the robot when the robot travels in a hairpin pipeline and L-shaped pipeline with a different
diameter pipe, respectively. In the experiment, the short radial curved holding mechanism with length of 40 mm
was used because of easy traveling in curved pipe. The supply pressure is 500 kPa. From Figure 14 (b), it was
found that the robot could travel in hairpin pipeline with total length of about 3 m within 240 s. From Figure 14(c), it was found that straight movements in T-shaped pipe were possible. It can be confirmed that the tested
robot can run through a pipe with a total length of 2.5 m in 150 s. In addition, it was also found that the7

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The 11th JFPS International Symposium on Fluid Power 2020

propulsive force could be generated when the EFPA bent further after contacting with the pipe. In the next
chapter, a novel propulsive method that utilizes this phenomenon as a radial bending type soft actuator will be
described.

(a) Schematic diagram of the complex pipeline pipe.

(b) Transient view of the movement of the robot in travelling in hairpin pipe line.

(c) Transient view of the movement of the robot in travelling in L-shaped pipeline
Figure 14. Results of travelling test using the tested robot in the complex pipe.

PROPULSION MECHANISM USING RADIAL BENDING TYPE SOFT ACTUATOR
Figure 15 shows the pipe propulsion mechanism using the radial bending type soft actuators by using the unique
phenomenon mentioned above. The propulsion mechanism consists of a back-to-back combination of two radial
bending type soft actuators with length of 40 mm, two shared pipeline plates and an omni wheels. It is noted that
there is no bending unit compared with the mechanism in Figure 11(a). In the shared pipeline, the same pressureand flow rate will be applied to three EFPAs in each holding mechanism. The size of the robot is 73 mm in outer
diameter, 200 mm in length, and 244 gf in weight.

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Figure 15. The pipe propulsion mechanism using the radial bending type soft actuators.

Figure 16 shows the transient view of the propulsion of the robot in the preliminary experiment. The operating
principle of the propulsion mechanism is as follows. At the beginning of movement of the propulsion
mechanism, the holding mechanism on the left side B does not hold the pipe, that is, the actuator is notpressurized. In the repeated operation from the second time, as shown in 1, the holding mechanism B is driven
and the robot repeats the operation from the state of holding the pipe. Next, as shown in 2, when the holding
mechanism on the right side A is driven, the robot holds the pipe wall with the tip of the EFPAs of the holding
mechanism by contacting with the pipe. Next, as shown in 3, when the EFPAs are further pressurized, an
entrainment operation of the EFPAs occurs at the contact point of the pipe wall as a fulcrum to give propulsion
to the propulsion mechanism. At this time, the holding mechanism B is contracted so as not to interfere with the
propulsive force generated by the holding mechanism A. Next, as shown in 4, EFPAs of the holding mechanism
B are pressurized, the tips of the holding mechanism B contacts with the pipe wall, and both holding
mechanisms hold the pipe. Next, as shown in 5 and 6, if the holding mechanism A is exhausted while theholding mechanism B is pressurized, the position of the holding mechanism A returns slightly. However, thisreverse propulsion operation is prevented by adjusting the holding force of the holding mechanism B. After that,as shown in 7, while the holding mechanism B is exhausted, the holding mechanism A is pressurized again, andboth holding mechanisms hold the pipe wall. Finally, as shown in 8, by exhausting the holding mechanism B
and further pressurizing the holding mechanism A, the force in the propulsion direction is generated to the
propulsion mechanism with the tip of the EFPA in contact with the pipe as a fulcrum, such as shown in 3. After
that, it returns to the operation 4 after the operation 8. The propulsion mechanism can travel in pipe by repeating
the operation from 4 to 8. The propulsion distance is the distance traveled by the operation from 2 to 3 or theoperation from 7 to 8. As a result, it can be confirmed that the tested propulsion mechanism can travel in thestraight pipe without sliding mechanism by only using entrainment operation of the EFPAs in the holding
mechanism. As a future work, it is necessary to investigate the optimal operating switching timing between
holding mechanism A and B.

Figure 16. The appearance of the operating principle of the robot.

CONCLUSIONSThis study aiming to develop the pipe inspection robot that can travel in various sized pipes by using only one
type pipe holding mechanism can be summarized as follows.
The bending soft actuator that the outer side of EFPA is restrained by connecting both ends through plastic
beam was proposed and tested. The radial bending type pipe holding mechanism using three bending actuators
based on EFPAs that can hold the pipe while bending radially was also proposed and tested. The holding

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characteristics using the prototyped holding mechanisms with a length of 40, 75 and 100 mm were investigated.
As a result of experiments for 75A, 100A and 150A standards pipes, the holding mechanism with a length of 75
mm has effective holding force for the 75A and 100A pipes. It was found that the holding force with a length of
40 mm was largest and effective.
The inchworm-type robot using two radial bending type pipe holding mechanisms with a length of 75 mm was
proposed and tested. The traveling test in different-diameter straight pipes of 75A, 100A, and 150A was carriedout. As a result, it was confirmed that the robot could travel in each pipe. It was also confirmed that the robot
using the holding mechanism with a length of 40 mm can travel in complex pipes including 75A and 100A.
The unique propulsion mechanism using only two radial bending type pipe holding mechanisms arranged backto-back was proposed and tested. As a result of the preliminary driving test, it could be confirmed that the tested
propulsion mechanism could travel in the straight pipe by only using entrainment operation of the EFPAs in
holding mechanisms.
As a future work, we are going to investigate the optimal operating switching timing between both holding
mechanisms in the tested propulsion mechanism.

ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Numbers 16K06202 and 19K04265.

REFERENCES
1. Current situation of water supply in japan (in Japanese), https://www.mhlw.go.jp/seisakunitsuite/bunya/topics/
bukyoku/kenkou/suido/newvision/chiikikondan/10/suishin_kondan10-1.pdf.
2. YUAN J., WU X., KANG Y., HUANG C., Development of an Inspection Robot for Long Distance Transmission
Pipeline on-site Overhaul, International Journal of Industrial Robot, Vol.36, 2009, p.546-550.
3. DERTIEN E., STRAMIGIOLI S., PULLES K., Development of an Inspection Robot for Small Diameter Gas
Distribution Mains, IEEE International Conference on Robotics and Automation, Shanghai, China, 2011, pp. 281-294.
SUZUKI, I., MATSUI, H., MATSUZAKA, D., Fundamentals of Water Hydraulic Control, New York Press, 1970, pp.123124.
4. NAGATANI, K., KUZE, M., YOSHIDA, K., Development of a Transformable Mobile Robot with a Variable Wheel
Diameter, Journal of Robotics and Mechatronics, Vol. 19, No.3, 2007, p. 252-257.
JOHNSON, H., et al., A Study of Performance Improvement of a Compressor, Transaction of the Japan Fluid Power System
Society, Vol.3, No.4, 2001, p.56-62.
5. YASUI, T., MANO, Y., ITO, F., NAKAMURA, T., Development of Joint Attachment for Passing Curved Pipe ofPeristaltic Motion Robot for In-pipe Inspection, Proc. of the 22nd International Conference on Climbing and Walking
Robots and the Support Technologies for Mobile Machines (CLAWAR), 2019, pp.45-52.
6. YAMAMOTO, T., KONYO, M., TADAKUMA, K., TADOKORO, S., High-speed Sliding-inchworm Motion
Mechanism with Expansion-type Pneumatic Hollow-shaft Actuators for In-pipe Inspections, Mechatronics, vol. 56, 2018,p. 101-114.
7. MIYASAKA, K., KAWANO, G., TSUKAGOSHI, H., Long-mover: Flexible Tube In-pipe Inspection Robot for LongDistance and Complex Piping, IEEE / ASME International Conference on Advanced Intelligent Mechatronics (AIM),
2018, pp. 1075-1080.
8. IKEUCHI, M., NAKAMURA, T., MATSUBARA, D., Development of an In-pipe Inspection Robot for Narrow Pipes
and Elbows using Pneumatic Artificial Muscles, IEEE / RSJ International Conference on Intelligent Robots and Systems,
2012, pp. 926-931.
9. KAWASAKI, T., AKAGI, T., DOHTA, S., HARIAHARA, K., Development of a Flexible Sliding Mechanism Using
Flexible Pneumatic Cylinder for Rescue Robot, FLUCOME 11th International Conference on Fluid Control,
Measurements, and Visualization, Keelung, Taiwan, 2011, 41.pdf, pp. 1-6.
10. AKAGI, T., DOHTA, S., WANG, L., KAWASAKI, T., Development of Sliding Type Inspection Robot using Flexible
Pneumatic Cylinder, International Journal of Advanced Mechatronic Systems, Vol. 5, 2013, pp. 87-94.
11. AKAGI, T., DOHTA, S., QIU, H., Pipe Inspection Robot using Flexible Pneumatic Cylinder Built-in Pneumatic Driving
System, International Forum on Systems and Mechatronics, Tainan, Taiwan, 2014, pp. 1-6.
12. CHOI, K., DOHTA, S., AKAGI, T., NINOMIYA, S., Development of Pipe Holding Mechanism for Pipe Inspection
Robot using Flexible Pneumatic Cylinder, International Conference on Mechanical, Manufacturing, Modeling and
Mechatronics, Kuala Lumpur, Malaysia, 2016, pp. 529-536.
13. KUSUNOSE, K., AKAGI, T., DOHTA, S., KOBAYASHI, W., NAKAGAWA, K., Development of Pipe Holding
Mechanism and Bending Unit using Extension Type Flexible Actuator for Flexible Pipe Inspection Robot, International
Journal of Mechanical Engineering and Robotics Research, Vol. 8, No. 1, 2019, p. 129-134.
14. HAYASHI, K., AKAGI, T., DOHTA, S., KOBAYASHI, W., SHINOHARA, T., KUSUNOSE, K., ALIFF, M.,
Improvement of Pipe Holding Mechanism and Inchworm Type Flexible Pipe Inspection Robot, International Journal of
Mechanical Engineering and Robotics Research, Vol. 9, No. 6, 2020, p. 894-899.

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15. KUSUNOSE, K., AKAGI, T., DOHTA, S., KOBAYASHI, W., SHINOHARA, T., HANE, Y., HAYASHI, K., ALIFF,
M., Development of Inchworm Type Pipe Inspection Robot using Extension Type Flexible Pneumatic Actuators,
International Journal of Automotive and Mechanical Engineering, Vol. 17, No. 2, 2020, p. 8019-8028.
16. SUZUKI, Y., AKAGI, T., DOHTA, S., KOBAYASHI, W., TAMAKI, H., MATSUI, Y., SHIMOOKA, S., Developmentof Tetrahedral Type Rehabilitation Device using Flexible Pneumatic Actuators, International Journal of MechanicalEngineering and Robotics Research, Vol. 7, No. 4, 2018, p. 409-414.
17. SHIMOOKA, S., DOHTA, S., AKAGI, T., KOBAYASHI, W., YONEDA, M., Development of Portable Rehabilitation
Device using Flexible Extension Type Soft Actuator with Built-in Small-Sized Quasi Servo Valve and Displacement
Sensor, JFPS International Journal of Fluid Power System, Vol. 11, No. 2, 2018, p.18-25.
18. SHIMOOKA, S., DOHTA, S., AKAGI, T., KOBAYASHI, W., SHINOHARA, T., Improvement of Home Portable
Rehabilitation Device for Upper-Limbs, JFPS International Journal of Fluid Power System, Vol. 12, 2019, No. 1, 2019, p.
10-18.

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Organized Session | Organized session

Soft Actuator 1[OS5-1-01] Analysis and Design of Servo Valve Using Buckled Tubes for Desired
Operation of Flexible Robot Arm〇Takumi Kobayashi1, Hideyuki Obayashi1, Tetsuya Akagi1, Shujiro Dohta1, Wataru Kobayashi1, Takashi Shinohara1, So Shimooka2 （1.Okayama University of Science, 2.Okayama
University）[OS5-1-02] Development of Pneumatic Variable Linear Stepping Actuator and Soft
Stepping Actuator with Bending Motion for Rehabilitation Device of Hip
Joint〇Kota Oe1, Tetsuya Akagi1, Shujiro Dohta1, Takashi Shinohara1, Wataru Kobayashi1, So
Shimooka2 （1.Okayama University of Science, 2.Okayama University）[OS5-1-03] Development of Extension Type Flexible Pneumatic Actuator with
Displacement Sensor Using Ring-shaped Magnet and Hall Sensor for
Tetrahedral-type Soft Mechanism〇Kenshiro Takeuchi1, Takumi Kobayashi1, Tetsuya Akagi1, Shujiro Dohta1, Takashi Shinohara1, Wataru Kobayashi1, So Shimooka2 （1.Okayama University of Science, 2.Okayama
University）[OS5-1-04] Development of Highly Durable Straight Fiber Type Pneumatic Artificial
Muscle with a Double Structural Air Chamber〇Naoki Saito1, Daisuke Furukawa1, Toshiyuki Satoh1, Norihiko Saga2 （1.Akita Prefectural
University, 2.Kwansei Gakuin University）[OS5-1-05] Evaluation of Lifting Motion with Non-wearing Type Pneumatic Power
Assist Device ~ Comparison of Active and Passive Type ~〇Masashi Yokota1, Reito Hirabayashi1, Masahiro Takaiwa2 （1.Tokushima University,
Graduate School Advanced Technology and Science, 2.Tokushima University, Graduate
School of Technology, Industrial and Social Sciences）©The Japan Fluid Power System Society

OS5-1-01

The 11th JFPS International Symposium on Fluid Power 2020

Analysis and Design of Servo Valve Using Buckled Tubes
for Desired Operation of Flexible Robot Arm
Takumi KOBAYASHI*, Hideyuki OBAYASHI *, Tetsuya AKAGI*, Shujiro DOHTA*,
Wataru KOBAYASHI*, Takashi SHINOHARA** and So SHIMOOKA ****Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science**
Science Dream lab., Okayama University of Science
1-1, Ridai-cho, Kita-ku, Okayama, 700-0005 Japan(E-mail: akagi@are.ous.ac. jp)***
Graduate School of Natural Science and Technology, Okayama University
3-1-1 Tsushima-naka, Kita-ku, Okayama 700-8530, Japan

Abstract. A Pneumatic soft actuator has many advantages of a higher ratio of generated force and lightweight andinexpensive. It is also attracting attention because of its intrinsic safety due to compliance based on air compressibility.
However, in a system using pneumatic soft actuators, a servo valve is most heavy and expensive. Therefore, in theprevious study, a low-cost and small-sized servo valve using buckled tubes was proposed and developed. In this paper,a static analytical model for designing the valve was proposed. In order to decrease the cost and size of valves, a multiport servo valve that can operate three pneumatic actuators of the flexible robot at the same time was proposed. The
design method of the valve based on the analytical model to get desired sinusoidal motion of the robot was described.
The driving test of the robot arm using the designed valve based on the model was carried out.
Keywords: Analytical model, Model based design, Servo valve using buckled tubes, Flexible robot arm, Extension
type flexible pneumatic actuator

INTRODUCTION
A driving system using pneumatic soft actuator have gained considerable attention and many studies have been
carried out [1-4]. They were developed as power assisting devices for nursing care [1], elderly [2] and worker [3]using pneumatic artificial muscles [4]. In this study, we aim to develop a low-cost home rehabilitation device
driven by pneumatic actuators. In a pneumatic driving wearable or portable device, the size and cost of a servo
valve becomes a serious issue. Typically, the most expensive and heavy control equipment in the pneumatic soft
mechanism is a servo valve. A typical electro-magnetic servo valve on the market has a complex mechanism formoving its spool while keeping a seal. The complex mechanism makes the valve bulkier and more expensive.
Therefore, a miniaturization and low-cost fabrication of a servo valve is a challenging issue. In order to decreasemass, size and cost of the valve, a simple mechanism for valve opening is required [5-10]. Zhao developed aquasi-servo valve using small-sized on/off valve [5]. Many researchers developed small-size valves driven by
vibration. One is a valve using a piezoelectric actuator [6]. The other is a valve using sound [7,8]. In the previous
study, a valve driven by a vibration motor was developed [9]. They were on/off type valves, not servo valves. Asa small-sized and low-cost servo valve, a servo valve using buckled tubes was proposed and tested in the previous
study [11-13]. The position control of the pneumatic rubber artificial muscle using the tested valve was carriedout [12]. A multi-port type servo valve was also developed to drive a double-acting type typical pneumatic
cylinder [13].
In the next step, to realize more compact valve, we consider developing a valve that can operate many actuators
at the same time so that the device using soft actuators can be work according to desired periodical motion. In thispaper, an analytical model of the servo valve to design the configuration of buckled tubes to get desired static
characteristics is proposed. As a driving target for design of the valve using the proposed model, we select a
flexible robot arm using three extension type flexible pneumatic actuators (we call it “EFPA” for short) developedin the previous study [14]. Because the motion of the tested robot arm is often required periodical rotational motion
as a rehabilitation device to give a passive exercise for arm and shoulder. In addition, to get the desired operationof three EFPAs in the flexible robot arm by using simple operation of the valve, multi-port control valve using
buckled tubes is designed based on proposed model. The driving test of the robot arm using the designed valve iscarried out.

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SERVO VALVE USING BUCKLED TUBES
Figure 1 shows the construction of multi-port type servo valve using buckled tubes to drive a double-acting typepneumatic actuator developed in the previous study [13]. The valve consists of four buckled tubes, a rotational
disk with two Y-shaped connectors (for outlet), four fixed straight connectors (for supply or exhaust port) and a
RC servo motor. Each buckled tube is connected between Y-shaped connector and fixed connector. Figure 2
shows the operating principle of the valve. By changing the rotational angle of the disk by motor, the buckled
angle of each tube is decreasing or increasing. By increasing the buckled angle of the tube, the sectional area ofthe tube increases to flow. In the opposite case when the buckled angle is decreasing, the sectional area in the tube
decreases and is closed surely below the threshold angle. Because the sectional area in the tube is related to therotational angle, the valve can realize to switch supply or exhaust for two outlets and control analog flow ratefrom both outlets.

Figure 1. Construction of the servo valve using buckled tubes for double-acting type pneumatic actuator [13].

Figure 2. Operating principle of the servo valve using buckled tubes for double-acting type pneumatic actuator [13].

Figure 3 shows the relationship between the motor rotational angle and output flow rate from two outlets of thetested valve. In Figure 3, the horizontal axis shows the incremental rotational angle of a rotation from the initial
angle of the motor (= 0 deg.), and the vertical axis shows the output flow rate from the left and right output ports.
The negative value shows exhaust flow rate, the positive value shows supply flow rate. The experimental resultsin Figure 3 show the average value of three repeated measurements. And the red and blue symbols show the output
flow rates from Port 1 and Port 2, respectively, which correspond to the left and right side ports in Figure 2. In theexperiment, the supply air pressure of 500 kPa was used. The discharge flow rate from each output port ismeasured with a digital flow meter (SMC Co., Ltd. PFMB7201-C8-AM). From Figure 3, it can be seen that theflow rates from both outlets can change in supply and exhaust symmetrically according to the rotational angle ofthe servo motor changes. It can also be confirmed that the valve has an overlap that the valve can maintain the
output pressure of the valve. The maximum flow rate of 50 l/min and more can be obtained from both ports. The
overlap area of the rotational angle is about ± 2 deg.

Figure 3. Relationship between motor rotational angle and output flow rate of the servo valve using buckled tubes.

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ANALYTICAL MODEL TO DESIGN THE VALVE
Figure 4 shows a simple analytical model of the servo valve using buckled tubes that is related to one side outlet
port based on geometrical configuration of tubes and connectors. The model consists of two buckled tubes for
supply and exhaust, a Y-shaped connector and two fixed straight connectors. In the model, the Y-shaped connector
can rotate around the origin O, a motor shaft, in Figure 4. By using the model, the location of the connected pointof tubes in Y-shaped connector after rotational motion can be calculated. As assuming the buckled tube keeps
straight except for the buckled point, the buckled point and buckled angle of each tube can also be calculated byusing law of cosines. In addition, by the investigated empirical formula of relation between buckled angle andsectional area of the tube described later, it can be calculated the static characteristics of flow rate according tomotor rotational angle for the initial buckled angle of each tube.

Figure 4. Model of the servo valve using buckled tubes.

From the geometric relationship, the bending angles θS and θE of two tubes change according to the rotation of
the rotating disk to which the Y-shaped connector is attached. If the initial angle from the motor rotational axis
(origin) to the central axis of the Y-shaped connector is θ0, the coordinates {XS, YS} and {XE, YE} of both tubes
connected points with the Y-shaped connector to which the tube is attached are given by the following equations.
𝑿𝑺 = −𝒓 𝐜𝐨𝐬(𝜽𝟎 − 𝜹𝜽)
𝒀𝑺 = 𝒓 𝐬𝐢𝐧(𝜽𝟎 − 𝜹𝜽)
𝑿𝑬 = 𝒓 𝐜𝐨𝐬(𝜽𝟎 + 𝜹𝜽)
𝒀𝑬 = 𝒓 𝐬𝐢𝐧(𝜽𝟎 + 𝜹𝜽)

(1)
(2)
(3)
(4)

where, r is the radius from the motor rotational axis to the end of Y-shaped connector, δθ is the motor rotational
angle, and subscript S and E show the supply and exhaust side buckled tube connectors, respectively.
From Equations (1) to (4), the distances between the Y-shaped connector and supply or exhaust connector L1TS
and L1TE are given by the following equation.
𝑳𝟏𝑻𝒊 = √(𝑿𝒊 − 𝒙𝒊 )𝟐 + (𝒀𝒊 − 𝒚𝒊 )𝟐 (𝒊 = 𝑺, 𝑬)(5)where, {xS, yS} and {xi, yE} are the coordinates of the center end of fixed supply and exhaust straight connectors,
respectively. In addition, as the supply and exhaust connectors are fixed, and these coordinate {xS, yS} and {xi, yE}
are constant. Therefore, from law of cosines, the bending angles θS and θE of the two buckled tubes can be obtainedby using Equation (5) and the distances L2T and L3T from each connector to the buckled point and are given by thefollowing equation.
𝜽𝒊 = 𝐜𝐨𝐬 −𝟏(𝑳𝟐𝑻 𝟐 + 𝑳𝟑𝑻 𝟐 − 𝑳𝟏𝑻𝒊 𝟐 )(𝒊 = 𝑺, 𝑬)
𝟐𝑳𝟐𝑻 𝑳𝟑𝑻(6)

The flow rates QS and QE from the buckled tubes on supply and exhaust side are given by the following equations.

𝑸𝑺 = 𝑨𝑺 ・𝑷𝑺 √𝟐𝑷𝑶・𝒇 ( )
𝑹𝑻
𝑷𝑺

(7)

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𝑸𝑬 = 𝑨𝑬 ・𝑷𝑶 √𝟐𝑷𝑨・𝒇 ( )
𝑹𝑻
𝑷𝑶(8)

where, AS and AE are the opening areas of the buckled tubes and are expressed as a function of the tube bending
angles θS and θE [rad]. PA, PO and PS indicate the atmospheric pressure, output pressure, and supply pressure,respectively, and R is the gas constant, T is the absolute temperature, and the function f (z) is the function
indicating the flow state, which is expressed by the following equation.
𝜿(𝒛𝟐⁄𝜿 − 𝒛(𝜿+𝟏)⁄𝜿 )(𝟎. 𝟓𝟐𝟖 ≤ 𝒛 ≤ 𝟏)
𝒇(𝒛) = √
𝜿−𝟏

(9)

𝜿
𝟐 𝟐⁄(𝜿−𝟏)(𝟎 ≤ 𝒛 ≤ 𝟎. 𝟓𝟐𝟖)()
𝜿+𝟏 𝜿+𝟏(10)

𝒇(𝐳) = √where, κ indicates the specific heat ratio. z means a pressure ratio between downstream and upstream.
In order to know the relationship of the opening area of the buckled tube with respect to the bending angle of thebuckled tube, the relationship of the flow rate with respect to the bending angle as shown in Figure 5 wasinvestigated. In the experiment, the same bending tube (SMC Co. Ltd, TUS0425) as the tube used in the tested
valve, the supply pressure of 500 kPa was applied. In the measurement, the discharge flow rate from the output
port to the atmosphere was measured with a digital flow meter (SMC Co. Ltd., PFMB7201-C8-A-M, resolution
1 l/min) according to the bending angle change of the tube. The bending angle of the tube was measured from thephotograph of the tube. In Figure 5, the vertical axis shows the mass flow rate calculated from the volumetric flowrate measured with the flow meter, and the horizontal axis shows the bending angle of the tube. In addition, fromthe repetition bending test of buckled tube, it has been confirmed that the durability of the buckled tubes is about
15,000 to 100,000 times under the critical condition of fast bending repetition. The durability of buckled tube isrelated to the generated heat by bending it. In typical use, we can also confirm that the valve using buckled tubes
can be used in more than three years.

Figure 5. Relation between angle of buckled tubes and mass flow rate.

Figure 6. Relation between buckled angle and opening area of buckled tube.

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From the result as shown in Figure 5, the opening area (sectional area) Ai of the tube was calculated by using
Equations (7), (8), and (10). Figure 6 shows the relation between the bending angle and the calculated sectional
area of the buckled tube. In Figure 6, the symbol shows the calculated value of the opening area obtained fromthe measured flow rate. The red line shows the result by using the approximate model described later (Equation(11)).
From the results as shown in Figure 6, the opening sectional area of the buckled tube with respect to the bending
angle is given by the following approximate equation.
𝑨𝒊 = 𝟎. 𝟖𝟒𝟓(𝟏 − 𝒆−𝟎.𝟏𝟎𝟓(𝜽𝒊−𝟔𝟖.𝟗) ) (𝒊 = 𝑺, 𝑬)(11)

Table 1 lists identified system parameters of the prototype valve. Using the proposed analytical model andidentified parameters mentioned above, it is possible to calculate the flow rate characteristics with respect to themotor rotational angle when the position of each connector and the length of each buckled tube are changed.
Table 1. Identified system parameters.

In order to confirm the effectiveness of the proposed model and the identified system parameters, the experimental
and calculated static characteristics of output flow rate were compared. Figure 7 shows the comparison between
the experimental and calculated results of the output flow rate with respect to the motor rotational angle. In thisfigure, the symbols ● and ■ indicate the measured flow rate from the outlet Port 1 and 2 of the valves, respectively.
In addition, ○ and □ show the calculated results based on the model. From Figure 7, the calculated overlap zone
within ± 3deg. is almost same as the experimental results that Port 1 has the overlap zone within ± 2deg. and Port
2 has the overlap zone from -4 to 3 deg. It can also be seen that the analytical model can predict the characteristics
of output flow rate relative well, although there is a slight difference in the maximum flow rate. It seems that it isbecause the relationship between the actual output flow rate (opening area) and the bending angle as shown in
Figure 6 has some gaps, it appears as an error. As a result, it can be confirmed the validity of the proposed modelthat can calculate from the motor rotational angle to the valve output flow rate can be confirmed. In addition, thedynamics of the valve depends on the dynamics of the RC servo motor with buckled tubes, that is a band width
frequency of 3.0 Hz.

Figure 7. Relation between motor rotational angle and output flow rate of the valve.

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DESIGN PROCEDURE OF MULTI-PORT SERVO VALVE AND TARGET DEVICE
As a design of the valve using the analytical model described above, we try to design a multi-port servo valve thatcan operate three pneumatic actuators at the same time with one RC servo motor. As a driving object using thedesigned valve, a flexible robot arm using extension type flexible pneumatic actuators (we call it “EFPA” forshort) was selected. Figure 8 shows the appearance and structure of EFPA used in the flexible robot arm. The
EFPA consists of the silicone rubber tube (inner diameter: 8 mm, outer diameter: 11 mm) is covered with a bellows
type sleeve (Swiftrans Co. Ltd., Stretching hose) to restrain its movement to the longitudinal direction. In EFPA,
the minimum and maximum diameter of the bellows are 12 and 20 mm, respectively. When the supply pressure
is applied to the EFPA, the silicone rubber tube expands in all directions, but the sleeve does not expand in thecircumferential direction and extends only toward the longitudinal direction. EFPA can extend until 2.5 times of
its original length. This large extension ratio is one of most important advantages of EFPA.

Figure 8. Appearance and schematic diagram of EFPA [14].

Figure 9 shows the appearance and construction of the flexible robot arm using EFPAs. The flexible robot arm
consists of 20 Y-shaped restraint PET plates and three reinforced EFPAs that is parallel arranged three EPPA
restrained with small restraint PET plates [14]. Three reinforced EFPAs set parallel with a radius of 70 mm every
120 deg. from the center axis by using Y-shaped restraint PET plates with thickness of 1 mm. The robot arm canbend toward radial direction by changing pressurized reinforced EFPA. The robot arm had been developed forrehabilitation device that can give passive exercise to patient’s arm and shoulder by its rotating motion whilebending. Therefore, the target motion of the robot arm is the rotating motion while bending.

Figure 9. Appearance and construction of reinforced EFPA with circumferential restraints [14].

As a specification of the designed valve, the valve needs three supply, exhaust and output ports. In addition, it isnecessary to realize the target motion of the robot arm by simply rotating the RC servo motor in the valve with arotational range of 60 deg., that is from 0 (neutral position) to -30 deg. (toward counterclockwise) and from 0 to
+30 deg (toward clockwise). By using this rotational motion of the motor, each port in the valve to drive each
EFPA so that the robot arm can rotate with the bending directional angle from 0 to 360 deg. while bending. The
design procedure to realize these specifications mentioned above is as follows.
(1st) Obtain the time-series data of the length of each EFPA based on the analytical model of the robot arm modelfrom the target movement of the robot arm.
(2nd) Match the calculated time-series data of each EFPA length to the time-series data of the rotational angle of
the RC servo motor.
(3rd) From time-series data of EFPA length, determine the switching timing from supply to exhaust and fromexhaust to supply of valve.

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The 11th JFPS International Symposium on Fluid Power 2020(4th) The expected initial bending angles of the supply and exhaust buckled tubes connected to each actuator are
decided. The calculated area for supply and exhaust using the decided initial bending angles and the model of thevalve draws on time-series data of EFPA length.
In this procedure, (4th) process is repeating until that the calculated area based on the model meets to the switching
timing determined by (3rd) process.

MODEL OF ROBOT ARM AND DESIGN OF MULTI-PORT SERVO VALVE
Next, as a design based on proposed models, the initial bending angle of the buckled tube of the multi-port servo
valve is calculated. First, according to the (1st) process in design procedure, the time-series data of the length ofeach EFPA with respect to the target trajectory is calculated by using the analytical model of the flexible robot
arm [14]. Figure 10 shows the analytical models of the robot arm. Figure 10 (a) shows the definition of the bending
directional angle α and the bending angle β of the robot arm, and Figure 10 (b) shows the definition of length ofeach EFPA, that is projected length of each EFPA on the plane according to the bending directional angle. The
reinforced EFPA located on the X axis is defined as EFPA 1. The other reinforced EFPAs arranged in acounterclockwise direction from the position of EFPA 1 are defined as EFPA 2 and 3, respectively. l1, l2, and l3
as shown in Figure 10 (b) are lengths of the reinforced EFPA 1, 2 and 3, respectively. Assuming that the shape ofeach EFPA in the robot arm is a circular arc, the lengths l1, l2, and l3 of each EFPA are given by the following
equations.
𝒍𝟏 = (𝑹 − 𝒓 ∙ 𝒄𝒐𝒔 𝜶) ∙ 𝜷
𝟐𝝅𝒍𝟐 = (𝑹 − 𝒓 ∙ 𝐜𝐨𝐬(− 𝜶)) ∙ 𝜷𝟑𝟒𝝅𝒍𝟑 = (𝑹 − 𝒓 ∙ 𝐜𝐨𝐬(− 𝜶)) ∙ 𝜷
𝟑
𝒍
𝑹=
𝜷

(12)
(13)
(14)
(15)

where, li (i = 1, 2, 3) is the length of each EFPA, R is the radius of curvature of the robot arm, r is the radius fromthe center to the circumference where the EFPA is located, respectively. From Equations (12) to (15), the central
length l of the robot arm is expressed by the following equation.
𝒍=𝒍𝟏 + 𝒍 𝟐 + 𝒍𝟑𝟑(a) Definition of angles α and β.

(16)(b) Definition of length of each EFPA.

Figure 10. Analytical model of the flexible robot arm [14].

In addition, to measure the bending directional angle α and bending angle β of the robot arm, an accelerometer(Kionix, Inc., KXM52-1050) was attached on the center tip of the robot arm. Both angles α and β obtained fromthat accelerometer output voltage can be obtained from the following equations.
𝜶 = 𝐜𝐨𝐬 −𝟏

𝑽𝒙
𝟐√𝑽𝒙 + 𝑽𝒚𝑽𝒛
𝜷 = 𝐜𝐨𝐬 −𝟏 (
)
𝑽𝒛𝒎𝒂𝒙

(17)
𝟐

(18)

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where, Vx, Vy, and Vz are the differential voltages from the initial values (horizontal state) of the accelerometer inthe x, y, and z directions, respectively. Vzmax is the maximum difference voltage of Vz when the center tip(accelerometer) is changed from horizontal and vertical.
Next, each desired length of EFPA for the target trajectory of the robot arm is calculated using the analyticalmodel of the robot arm. Here, the target trajectory was set so that the bending angle β was 120 deg. (2π / 3) andthe bending directional angle α was 0 to 360 deg. (0 to 2π) in 60 seconds. Figures 11 (a), (b) and (c) show thecalculation results of the time-series changes in the displacement of EFPA 1, 2 and 3 with respect to this target
trajectory, respectively. In the calculation (experiment), the natural length of each EFPA constructing the robot
arm was set to 230 mm under the condition when each EFPA was arranged at a radius of 70 mm from the centerof the robot arm.

(a) EFPA 1(b) EFPA 2(c) EFPA 3
Figure 11. Transient response of calculated length of each EFPA using the analysis model.

According to the (2nd) process in the design procedure, the time series data of the calculated lengths matches tothe time series data of the rotational angle of the RC servo motor (0 → -30 → 0 → 30 → 0 deg.). Figure 12 shows
data set of the desired length of each EFPA and rotational angle of the RC servo motor.
According to (3rd) process in design procedure, and valve switching timing for supply or exhaust is decided. In
detail, the timing at maximum length of each EFPA is regarded as a switching timing from supply to exhaust. The
timing of minimum length of each EFPA, it is also regarded as a switching timing from exhaust to supply. Because
the motion of EFPA while exhausting is slower than the motion while supply, the switching timing from exhaustto supply is also changed a little later. In (4th) process, by paying attention to these switching points, the initial
bending angle of both supply and exhaust buckled tubes so that supply and exhaust motion area can agree withthe desired switching timing by using the model of valve mentioned above.
By repeating (4th) process, the initial bending angle of each valve was determined. In Figure 12, the red, blue andgreen broken lines indicate the supply, exhaust and holding area that are calculated by the designed initial bending
angle of each buckled tube, respectively. In addition, designed initial bending angle for each valve is also indicatedon right side of each figure. From Figure 12, it can be seen that the calculated valve operation based on the modeland the determined initial bending angle almost meet the desired switching timing, although there is some areathat cannot be realized due to restrictions on repetitive motor operation.

Figure 12. Data set of the desired length of each EFPA and rotational angle of the RC servo motor.

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Figures 13 (a), (b) and (c) show the calculated static relations between the motor rotational angle and the output
flow rate of each designed valve connected to EFPA 1, 2 and 3 using the analytical model of the valve and designed
initial bending angles of buckled tubes, respectively. From these results, it can be seen that the output flow rate of
the valve for EFPA 1 is inverse to others. It is found that the overlap zone of each valve is different, that are
overlap from -9 to 3 deg. for EFPA 1, the overlap from -3 to 3 deg. for EFPA 2 and the overlap from -9 to 15 deg.
for EFPA 3, respectively.

(a) Valve for EFPA 1(b)Valve for EFPA 2(c)Valve for EFPA 3
Figure 13. Calculated relations between the motor rotational angle and the output flow rate of each designed valve.

DRIVING TEST OF ROBOT ARM USING THE DESIGNED VALVE
Figure 14 show the tested multi-port servo valve using buckled tubes based on design using the proposed model.
The valve has three 3-port type valve units that each valve unit consists of two buckled tubes for supply andexhaust. These are arranged in parallel. Three rotational disks of valve units are driven by a RC servo motor at
the same time. Then, the designed initial buckled angle for each valve is shown in Figure 12. The tested valve
also has a rotational potentiometer to monitor the rotational angle of the motor. The size of the valve becomescompact compared with the amount volume of three original 3-port type valves. The valve has the length of 90
mm, the width of 81mm and height of 105 mm.

Figure 14. Tested multi-port servo valve using buckled tubes.

Figures 15 shows the schematic diagram and appearance of a control system using the designed multi-port servo
valve and the flexible robot arm. The control system consists of the flexible robot arm using three reinforced
EFPAs, the designed multi-port servo valve, an accelerometer (Kionix, Inc., KXM52-1050) and an embedded
controller (Renesas Electronics Co. Ltd., H8 / 3664F). The accelerometer is attached to the center tip of the robot
arm and is used to measure the bending directional angle α and the bending angle β of the arm. The RC servo
motor of the valve is operated by changing the duty ratio with respect to the preset target angle via the PWM port
in the embedded controller. The output signals from the accelerometer are monitored via a recorder (Graphtec
Corporation, GL900). In the experiment, the supply pressure of 300 kPa was applied to the multi-port servo valve.
The RC servo motor in the valve was repeated to rotate from 0 to -30 to 0 to +30 to 0 deg. every 1 deg. per second.

Figure 15. Schematic diagram and appearance of the control system using the designed valve and the flexible robot arm.

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Figure16 shows the transient view of motion of the flexible robot arm using the tested valve. It can be confirmedthat the robot arm can move according to the desired motion by giving the swing motion of RC servo motor as
shown on the lower right of Figure 12.

Figure 16. Transient view of motion of the flexible robot arm using the tested valve.

Figures 17 (a) and (b) show the transient response of the measured bending directional angle α and bending angle
β of the flexible robot arm by using the accelerometer when the robot arm moves according to sequential operation
using the tested valve, respectively. Figure 18 shows the transient response of the measured rotational angel ofthe RC servo motor by using the potentiometer attached on the valve. In both results, a low-pass filter with 1.5
Hz was used to avoid noise. In Figure 17, the solid red line shows the experimental result, and the broken blueline shows the desired bending directional angle and bending angle. From Figure 17, it can be confirmed that the
robot arm can almost keep constant bending angle of 120 deg., that is the desired bending angle. It can also trace
the desired bending directional angle with the rotation from 0 to 360 deg. It seems that the error is caused by thesequential control without feedback control. In addition, by adjusting the swing speed (rotational speed) of RC
servo motor, we believe that the tracking error will be improved. As a result, it can be confirmed that the developed
valve can operate three EFPAs in the robot arm at the same time with simple repeating rotational motion using
one RC servo motor, and the robot arm can roughly trace the target orbit.

(a) Bending directional angle α.

(b) Bending angle β.

Figure 17. Transient response of the measured bending directional angle α and the bending angle β of the robot arm.

Figure 18. Transient response of the measured rotational angel of the RC servo motor.

CONCLUSIONSIn this paper, the analytical model of the servo valve using buckled tubes to design the configuration of buckled
tubes and to get the desired static characteristics was proposed. The flexible robot arm using parallel arranged
three EFPAs was also tested. In order to get the desired motion of the robot arm by giving simple swing motionof the valve, the initial bending angle of buckled tubes for each port of the valve was designed based on the
proposed model. And the multi-port servo valve using buckled tubes with the designed initial bending angle wasdeveloped. The driving test using the tested valve was carried out. As a result, the robot arm could move according
to the desired motion. It can be confirmed that the proposed model is valid to design the valve using buckled tube
and is useful to design the valve with the special purpose of use.

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ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Numbers 16K06202 and 19K04265.

REFERENCES
1.
2.
3.
4.
5.
6.
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ISHII, M., YAMAMOTO, K., HYODO, K., Stand-Alone Wearable Power Assist Suit -Development and Avail-ability-,
Journal of Robotics and Mechatronics, Vol.17, 2005, pp.575-583.
NORITSUGU, T., TAKAIWA, M., SASAKI, D., Development of Power Assist Wear Using Pneumatic Rubber
Artificial Muscles, Journal of Robotics and Mechatronics, Vol.21, 2009, pp.607-613.
KOBAYASHI, H., ISHIDA, Y., HASHIMOTO, T., Muscle Suit Development and Factory Application, International
Journal of Automation Technology, Vol.3, No.6, 2009, pp.709-715.
NAGATA, Y. ed., Soft Actuators -Forefront of Development-, Tokyo: NTS Ltd., 2004, pp.291-335.
ZHAO, F., DOHTA, S., AKAGI, T., Development and Analysis of Small-Sized Quasi-Servo Valve for Flexible Bending
Actuator, Transactions of JSME, Series C, Vol.76, No.772, 2010, pp.3665-3671.
JIEN, S., HIRAI, S., HONDA, K., Miniaturization Design of Piezoelectric Vibration-Driven Pneumatic Unconstrained
Valves, Journal of Robotics and Mechatronics, Vol.22, No.1, 2010, pp.91-99.
JING, S. et al., A Study on a Sound Operated Valve for a Wearable Pneumatic System: 1st Report: Development of a
Pilot Valve a Sound Operated Direction Control Valve, Journal of the Japan Fluid Power System Society, 39(3), 2008,
pp.54-60 (in Japanese).
NISHIOKA, Y. et al., A New Pneumatic Control System using Multiplex Pneumatic Transmission, Proceedings of 7th
JFPS International Symposium on Fluid Power Toyama, 2008, pp.439-442.
AKAGI, T., DOHTA, S., KATAYAMA, S., Development of Small-sized Flexible Pneumatic Valve using Vibration
Motor and Its Application for Wearable Actuator, Journal of Computer Applications in Technology, 39(1/2/3), 2010,
pp.86-92.
AKAGI, T., DOHTA, S., Development of Wearable Pneumatic Actuator and Multiport Pressure Control Valve, Journalof Robotics and Mechatronics, 17(5), 2005, pp.529-536.
NASIR, A. et al., Development of Low-Cost Wearable Servo Valve Using Buckled Tube, Applied Mechanics andMaterials, 393, 2013, pp. 532-537.
NASIR, A. et al., Development of Small-sized Servo Valve Controlled by Using Buckled Tube and Its Application,
Journals of the Japan Society of Mechanical Engineers, 7(4), 2013, pp. 516-527.
OBAYASHI, H. et al., Development of 5–Port Type Low–Cost Servo Valve Using Buckled Tubes, International Journalof Mechanical Engineering and Robotics Research, Vol. 7, No. 6, 2018, pp. 669-673.
SHIMOOKA, S. et al., Development of Reinforced Extension Type Flexible Pneumatic Actuator with Circumferential
Restraints and Its Application for Rehabilitation Device, International Journal of Automotive and MechanicalEngineering, 17(3), 2020, 8116-8127.

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Development of Pneumatic Variable Linear Stepping Actuator andSoft Stepping Actuator with Bending Motion for Rehabilitation
Device of Hip Joint
Kota OE*, Tetsuya AKAGI*, Shujiro DOHTA*, Takashi SHINOHARA**,
Wataru KOBAYASHI* and So SHIMOOKA****

Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science**
Science Dream lab., Okayama University of Science
1-1,Ridai-cho, Kita-ku, Okayama, 700-0005 Japan(E-mail: akagi@are.ous.ac. jp)***
Graduate School of Natural Science and Technology, Okayama University
3-1-1 Tsushima-naka, Kita-ku, Okayama 700-8530, Japan

Abstract. Based on elderly society in Japan, the rehabilitation device that can be used in home or hospital withoutcare person’s assistance has been desired. In the device, a soft actuator that can move with both longer strokes andlarger force is useful. In the previous study, a Pneumatic Linear Stepping Actuator (“PLSA” for short) that can pushand pull the flexible rod while changing the gripping position of the tube was developed. In this study, a variable step
type PLSA is proposed and tested. And the development of a rehabilitation device for hip joint that using three PLSAs
and one connecter connected with three flexible rods and a knee supporter. The tracking position control of knee
supporter based on an analytical model is also described. To make simpler rehabilitation device, a soft stepping
actuator with bending motion is also proposed and tested. The control system using developed actuators is also
introduced.
Keywords: Pneumatic variable linear stepping actuator, Soft stepping actuator with bending motion, Rehabilitation
device for hip joint, Backdrivability, Passive exercise.

INTRODUCTION
In 2020, the ratio of the Japanese elderly becomes more than 26 % [1]. In 2040, the ratio will be expected about
36%. Under such a critical society condition, various pneumatic drive soft mechanisms to assist welfare work forthe elderly and disabled[2] have been actively developed[3-10]. As a pneumatic soft actuator, a McKibben typepneumatic artificial muscle (PAM for short) is well known and used for soft mechanism. The PAM can generatelarger force of more than 300 N by contracting toward the longitudinal direction. However, the contract ratio isvery small, that is less than 25 % of the original length. As a soft mechanism using PAM, T. Noritsugu and H.
Kobayashi developed power assisted wears[3-6] to increase the force for welfare workers and heavy labors. Asan application using other pneumatic soft actuator, H. Taniguchi and H. Kawasaki developed a rehabilitation
device for inhibiting contracture of finger joints [7,8]. As a soft mechanism using rigid pneumatic actuator, M.
Takaiwa also developed a rehabilitation device for contracture patients using pneumatic parallel manipulator byrealizing softness using controller[9]. In these mechanisms, a pneumatic soft actuator was often used because it
has an advantage of compliance based on air compressibility. However, most of soft actuator can not realize longstroke motion because of its flexibility. Based on the issue, in the previous study, the Pneumatic Linear Stepping
Actuator (we call it “PLSA” for short) that can realize both large force of more than 300 N and long stroke motion
was proposed and tested[10,11]. The tested actuator can push or pull the flexible rod with a certain step bychanging the gripping point of the flexible rod. It means that the stroke of the actuator has no limit, it depends onlength of the flexible rod. The flexible robot arm using three PLSAs that can extend like a robot of Dr. Octopus
in cartoon “Spider-man” was also developed [12]. The holding mechanism in PLAS was improved so as to holdthe flexible rod surely[13]. In addition, a simple rehabilitation device using PLSA for upper limb was also
developed[14]. However, the tested PLSA moves only every certain stroke. Therefore, the positioning accuracy
depends on its stroke.
In this study, a novel pneumatic linear stepping actuator that can change the stroke for one step freely is proposed
and tested. The construction and operating principle of the tested actuator are described in this paper. As an

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application using proposed these actuators, rehabilitation devices using three previous PLSAs and renewed PLSAs
for hip joint are also described. The control method to improve positioning accuracy was also introduced.

PREVIOUS PNEUMATIC LINEAR STEPPING ACTUATOR
Figure 1 shows the construction of the previous PLSA. The PLSA consists of three double acting type pneumatic
cylinders with the stroke of 50 mm and the inner diameter of 16 mm (Koganei Co., PBDA 16×50-M), twopneumatic chucks and a flexible soft polyurethane tube with the outer diameter of 12 mm and the inner diameterof 10 mm as a rod. The ends of three cylinders are connected with a base stage with the pneumatic chuck. They
are set every 120 degrees with radius of 25 mm from the center of the base stage. The rod ends of three pneumatic
cylinders are also connected with the moving stage with the pneumatic chuck. The flexible linear stepping actuator
has a length of 150 mm and an outer diameter of 68 mm. The mass of the actuator is 0.7 kg. The actuator canrealize a flexibility at the point of working rod end. The actuator also has a backdrivability while the pneumaticchuck is not driven.

Figure 1. View and schematic diagram of the previous pneumatic linear stepping actuator.

Figure 2 shows the schematic diagram of a magnetic core type pneumatic chuck developed in our previous
study[13]. The pneumatic chuck consists of two ring-shaped neodymium magnets, I-shaped cylindrical plastic
core in the flexible tube, a doughnut-shaped diaphragm (that is made of a rubber film with a thickness of 1 mm,
an inner diameter of 24 mm and an outer diameter of 36 mm) and a mechanical chuck. The mechanical chuckconsists of two claws, fixed and moving bases with a slope angle of 45 degrees. The cylindrical plastic core
consists of ring-shaped plastic disks and a ring-shaped neodymium magnet with the outer diameter of 8.5 mm andthe inner diameter of 5.5 mm. They are penetrated by a screw and a nut. In the mechanical chuck, the ring-shaped
neodymium magnet with the outer diameter of 20 mm and the inner diameter of 12 mm is set on the flexible tube
to attract the magnet of the moving core. By attracting both magnets of the chuck and the core each other, the core
can automatically track the movement of the chuck. When the supplied pressure is applied to the diaphragm, theflexible tube is clamped from both inner and outer sides by claws and moving core. By this method, the chuck
can hold the tube (rod) surely.

Figure 2. Schematic diagram of inner construction of magnetic core type chuck.

Figure 3 shows the operating principle of the PLSA. The operating principle is as follows: First, the base sidechuck is driven to hold the tube as shown in Figure 3 (1). In this condition, as shown in Figure 3 (2), the rightside chambers of cylinders are pressurized. A moving stage with the pneumatic chuck moves toward right. Whenthe moving stage reaches at the right side stroke end, the chuck on the stage is activated to hold the tube as shownin Figure 3 (3). After that, as shown in Figure 3 (4), the chuck on the moving stage is driven and the chuck on the2

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The 11th JFPS International Symposium on Fluid Power 2020base stage is released. Under the condition that the chuck on the moving stage holds the flexible tube, the air in
the right side cylinder chamber is exhausted. At the same time, as shown in Figure 3 (5), the left side chambers of
the cylinders are pressurized. While the chuck on the moving stage keeps holding the tube, the flexible tube ispushed toward left as shown in Figure 3 (6). By repeating this procedure, the tube can move toward left every
certain stroke. As a result of measurement of generated force, the maximum generated force of 260 N could beobtained.

Figure 3. Operating principle of the flexible linear stepping actuator.

REHABILITAION DEVICE FOR HIP JOINT USING PLSA
As an application of PLSA, we select a rehabilitation device for hip joint. The reason why our target is to develop
a rehabilitation device for hip joint is as follows. Now the rehabilitation device for hip joint on the market is a
device that a patient drives it by him/herself. Therefore, it seems that the device that can give passive exercise for
hip joint is useful as a voluntary rehabilitation. In the typical rehabilitation of a hip joint, a Physical Therapist(“PT” for short) moves a patient’s knee so that its moving orbit can draw ellipse. By this method, the PT gives apassive exercise to the patient’s hip joint. By taking account of moving orbit of the knee, a prototype rehabilitation
device for hip joint using PLSA is proposed and tested. Figure 4 shows the construction of the prototype
rehabilitation device of hip joint using three PLSAs. The device consists of three PLSAs with flexible rods, a knee
supporter and an experimental base frame. Three PLSAs are set on the base frame so that each PLSA is locatedevery 120 degrees from the center, that is the initial position of knee supporter as shown in Figure 4. Each flexible
rod is connected to each corner of a triangle-shaped moving stage. The knee supporter is connected with themoving stage by a screw. The operating principle of the tested device is as follows. By changing rod length fromeach PLSA at the same time, the position of the knee supporter can be changed. In addition, as the rod is flexible,
the position of the knee supporter can be controlled while each rod is bending. However, to control the knee
position, the bending rod becomes uncertain element. In addition, in the case when the rod bends sharply at theend of mechanical chuck, it was found that the frictional force becomes rapidly large.

Figure 4. Construction of the prototype rehabilitation device for hip joint using 3 PLSAs.

Therefore, to decrease uncertainness based on bending rods and affection of friction, a universal-joint-type stagefor PLSA is proposed and tested. Figure 5 shows the construction of the universal stage with PLSA. The universal
stage consists of a rotational type universal joint with radial bearings and a wire-type linear potentiometer[15].
The universal stage can rotate around pitch and yaw axes. Several small balls between lower base and stage work

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as a thrust bearing. The wire-type linear potentiometer is used to measure the rod length between the PLSA andthe moving stage with knee supporter.

Figure 5. View and schematic diagram of the universal stage with PLSA.

Figure 6 (a) and (b) show the appearance of improved rehabilitation device for hip joint using PLSAs with
universal stages and the schematic diagram of control system using three PLSAs and an embedded controller,
respectively. In the improved rehabilitation device, the rigid aluminum pipe is installed into the flexible rod ineach PLSA, because the universal stage adjusts the suitable direction of PLSA naturally. The aluminum pipe withthe outer diameter of 8 mm works as an inner core instead of the moving core. The mechanical chuck grasps therod surely. The control system of rehabilitation device as shown in Figure 6 (b) consists of 12 on/off valves(Koganei Co. Ltd., G010E-1), the embedded controller (Renesas electronics Co. Ltd., SH7125) and the wire-typelinear potentiometer. Each PLSA is driven by four on/off valves. Two valves are used to drive three pneumatic
cylinders for pushing and pulling. Each mechanical chuck is driven by an independent on/off valve. Each on/off
valve is driven through transistor and I/O ports of the embedded controller. The end of the wire connected to thestage with the knee supporter in order to measure the rod length through A/D converter for monitoring.

(a) the improved rehabilitation device for hip joint.

(b) Schematic diagram of control system.

Figure 6. Appearance of the improved rehabilitation device for hip joint using PLSAs with universal stages and theschematic diagram of control system.

ANALYTICAL MODEL AND CONTROL BASED ON STEPPING MOTIONSIn order to control the position of knee supporter by controlling the length of three rods of PLSAs, it needs ananalytical model that can calculate rod lengths of three PLSA 1, 2 and 3 from the desired position of the knee
supporter. Figure 7 shows the simple analytical model of the rehabilitation device for hip joint. In the model, theposition of the knee supporter can be obtained as an intersection of three circles from each center of PLSA. The
initial position of knee supporter is defined as an origin of moving area. Three PLSAs, that is Actuator 1, 2 and 3,
were located on the coordinate (0, -b, 0), (a, b, 0) and (-a, b, 0), respectively.

Figure 7. Simple analytical model of the rehabilitation device for hip joint.

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From geometric relations in the model, the length of each rod, that is each radius r1, r2 and r3, can be calculatedby the following equation.
𝑟1 = √𝑥 2 + (𝑦 + 𝑏)2 + 𝑧 2(1)

𝑟2 = √(𝑥 − 𝑎)2 + (𝑦 − 𝑏)2 + 𝑧 2(2)

𝑟3 = √(𝑥 + 𝑎)2 + (𝑦 − 𝑏)2 + 𝑧 2(3)

where (x, y, z) means the desired coordinate. a and b show the distance from the origin to the center of PLSA
toward x and y direction, respectively. In the experiment, a of 48 mm and b of 38 mm were used. The position
control of the knee supporter is carried out as follows. First, the embedded controller calculates the desired lengthof each rod (radius) from the desired position of the knee that had installed into memory on the embedded
controller based on the model mentioned above. The controller also calculates the error ei between the desired
length ri and the resent length of each rod Li(k). Based on the following equation, each PLSA is driven based onerror.
（𝑖 = 1,2,3）

𝑒𝑖 = 𝑟𝑖 − 𝐿𝑖𝑆𝑒𝑖 ≥ 𝑡 :2𝑒𝑖 ≤ −
−

𝑆𝑡
2

𝑆𝑡
2

:

≤ 𝑒𝑖 ≤

𝑆𝑡
2(4)

𝑛𝑖 = 1（Push）(5)𝑛𝑖 = −1(Pull)(6)(hold)(7): 𝑛𝑖 = 0

where St and ni mean the stroke of the cylinder (that is the displacement of 50 mm for one step motion), a motion
index for push(= 1), pull(= -1) or hold(= 0), respectively. In the control, each PLSA is selected “push”, “pulls” or“hold” based on error is within a half of one step motion of 50 mm. Then, the current length of each rod Li can beobtained by the integral of step motion based on the motion index ni, and given by the following equation.
𝐿𝑖 = 𝐿𝑖0 + 𝑆𝑡 𝛴𝑛𝑖(8)where Li0 shows the initial length of each PLSA. By utilizing the characteristics of the stepping actuator, thecontroller can calculate the current length of each PLSA from the previous motion of each PLSA without senor.
In order to drive three PLSA at the same time based on each motion index, 27 set of driving patterns for 12 on/off
valves were prepared in the embedded controller for the position control of the device. Figure 8 shows a sampleof time chart in the case when Actuator 1 is pushing (n1 = 1), Actuator 2 is pulling (n2 = -1) and Actuator 3 isholding (n3=0). The Valves 1 to 4 in Figure 8 were shown in Figure 6 (b). The Valves 5 to 8 were used forActuator 2 and Valves 9 to 12 were for Actuator 3. All driving patterns is carried out every 2.6 s. Figure 9 shows
the transient view of movement of the knee supporter while the position of the knee supporter is tracing to an
elliptical target trajectory with the major axis length of 400 mm and the miner axis length of 320 mm. From Figure
9, it can be seen that the position of the knee supporter can trace the elliptical target trajectory relatively well.
Actueator

1

times
Valve 1

0.5s

1.0s

1.5s

1.8s
1.3s

Valve 51.8s

Valve 6

1.3s

Valve 7

1.3s

0.5s

Valve 8
Valve 9
3

3.0s

1.3s

Valve 4

2

2.5s
0.5s

Valve 2
Valve 3

2.0s

1.3s

1.3s
2.6s

Valve 101.8s

Valve 11
Valve 122.6s

Figure 8. Time chart of pattern(n1=1, n2=-1, n3=0).

Figure 9. Transient view of movement of knee supporter in control.

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CONTROL SCHEME FOR VARIABLE LINEAR STEPPING ACTUATOR
However, the PLSA pushes and pulls every a certain stoke of 50 mm. In other words, the position of the knee
supporter has always error within 50 mm. In addition, in the case when a slipping of the rod occurs at themechanical chuck by some disturbance, the error becomes larger. Therefore, it needs feedback control by using
displacement senor. Based on this issue, the improved PLSA with universal stage equipped wire-type linear
potentiometer. However, even if the feedback controller is equipped in the system, it still has a problem that the
PLSA moves every 50 mm because it is a stepping actuator. Therefore, the method to adjust the step stroke
arbitrarily using PLSA is proposed. Namely, we tried the method that both mechanical chucks are used as a brake.
The proposed method is as follows. During the pushing and pulling operation of PLSA, either mechanical chuckholds the rod certainly. In the other word, the opposite mechanical chuck is available to use as a brake. In detail,in the pushing operation, the opposite mechanical chuck is driven when the error becomes positive to negative.
To realize the operation, the embedded controller detects the rod length through A/D converter and the linear
potentiometer every 0.5 s. In the opposite case when pulling, the chuck is driven when the error becomes negative
to positive.
Figure 10 (a), (b) and (c) show the transient responses of the rod length of Actuator 1, 2 and 3 when the
rehabilitation device moves the knee supporter according to the elliptical target trajectory, respectively. In Figure
10, green triangles and red circles show the results using the previous and proposed control scheme. The bluelines show the desired length. It can be seen that the controlled length using the proposed method can trace thedesired length well. The standard deviation of tracking control error using the previous and proposed methods are
19.1 mm and 4.7 mm, respectively.

(a) Actuator 1(b)Actuator 2(c) Actuato r 3
Figure 10. Transient response of each rod length in 3 PLSA when the rehabilitation device moves the knee supporter
according to the elliptical target trajectory.

SOFT STEPPING ACTUATOR WITH BENDING MOTION
The rehabilitation device for hip joint using PLSA was successfully developed mentioned above. However, fromview of home-based rehabilitation device, the device requires larger setting area and moving area, because theopposite side rod moves according to the motion of knee supporter. In ideal, a compact rehabilitation device thatcan be used by hanging is necessary. As a hanging type rehabilitation device, the device needs the following
functions.
1) Function to adjust the hanging position of knee supporter.
2) Function to rotate the knee supporter according to target orbit.
In order to satisfy both specifications, the actuator needs to adjust the longitudinal length of rod. In addition, theactuator must bend toward radial direction to realize the rotational motion of knee supporter, that is to require soft
mechanism. In this study, in order to reconstruct the ordinal PLSA to soft mechanism, we use an Extension type
Flexible Pneumatic Actuator (we call it “EFPA” for short) instead of the rigid pneumatic cylinder. Figure 11
shows the view and schematic diagram of EFPA [16,17].

Figure 11. View and schematic diagram of EFPA[16,17].

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The EFPA consists of a silicone rubber tube with inner diameter of 8 mm, outer diameter of 11 mm and covered
with a bellows type nylon sleeve (The Fit Life Co. Ltd., Hose Reel). The EFPA can extend about 2.5 times of its
original length. The maximum pulling force of the EFPA is about 60 N[16,17]. The EFPA is very lightweight, themass is 50 g for the original length of 200 mm. The material cost of EFPA is low, it is about 5 US dollars per 1m.
In order to construct the sliding mechanism such as a parallel arranged pneumatic cylinders, a cross arranged
reinforced EFPA is proposed and tested. Figure 12 shows the construction of the improved PLSA with the crossarranged reinforced EFPA. The improved PLSA consists of the sliding mechanism using the reinforced EFPA,
two mechanical chucks and the flexible rod. The cross reinforced EFPA consist of six EFPAs, seven restrain PET
sheets and two base connectors. In the cross reinforced EFPA, each three EFPAs are used to construct a
tetrahedral-shaped actuator. And two tetrahedral-shaped actuators are set so as to unite both actuators face to facewith rotational difference of 60 degrees from the center. Both side of six EFPAs are connected to two baseconnectors. By this cross reinforced arrangement, the reinforced EFPA can push and pull according to the center
axis, the PLSA using the soft actuators can be realized.

(a) Initial state(b) Pressurized(500kPa)
Figure 12. View of the improved PLAS with cross reinforced EFPAs.

In order to realize a bending operation for rotating the knee supporter according to the target orbit, a soft actuatorthat can change the bending direction of the flexible rod while pushing and pulling the rod is proposed and tested.
Figure 13 (a) shows the construction of a soft stepping actuator with bending motion. The soft stepping actuatorconsists of a bending unit and the improved PLSA with cross reinforced EFPAs. The bending unit and theimproved PLSA are connected in serial. The bending unit consists of six EFPAs with a natural length of 110 mm,
ten restrain PET sheets and two donut-shaped end connectors. In the bending unit, six EFPAs are set parallel on
the position with radius of 40 mm every 60 degrees from the center, the flexible rod passes through the center
hole in restrain PET sheets. The tested soft stepping actuator with bending motion has the outer diameter of 80
mm and the natural length of 380 mm without the flexible rod. The mass of the soft stepping actuator is about 1.8
kg. Figure 13 (b) shows the schematic diagram of control system of the tested actuator. The control system consistsof the tested stepping actuator, ten on/off valves and an embedded controller. In the soft stepping actuator, theimproved PLSA is driven by four on/off valves, that are Valve 7 to 10, and the bending unit is driven by six on/off
valves, that are Valve 1 to 6. The Improved PLSA pushes and pulls the flexible rod toward the bending unit. The
bending unit works to change the direction of the pushing or pulling rod. The bending unit can bend toward 12
radial directions by driving three or four adjacent on/off valves.

(a) Construction of soft stepping actuator(b) Schematic diagram of control system
Figure 13. Construction of soft stepping actuator with bending motion and schematic diagram of control system.

In order to realize similar operation using the rehabilitation device for hip joint using PLASs, the preliminary
experiment using the tested soft stepping actuator with bending motion was carried out. In the experiment, thetested actuator was set vertically on base frame so that the bending unit set toward lower. The end flexible rod
connected with the moving stage with the knee supporter. Figure 14 shows the transient view of movement of the7

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tested actuator with knee supporter while the actuator moves the knee supporter according circular motion. In
detail, after adjusting the length of rod by using the improved PLSA, the bending unit bends so that the bending
directional angle of the unit can change from 0 to 360 degrees every 30 degrees per 2 s. From Figure 14, it can beseen that the tested actuator can realize the rotational motion of knee supporter so as to trace a desired orbit, even
if the more compact size of the device is used compared with the case using the previous one as shown in Figure
9. In addition, as the tested soft stepping actuator can adjust the pushing, pulling force and bending force bychanging supply pressure to EFPAs, it seems that the tested actuator can apply to other rehabilitation device thatis necessary to adjust the applied force to a patient as a future work.

Figure 14. Transient view of movement of tested soft stepping actuator while it moves the knee supporter according
circular motion.

CONCLUSIONSIn this study, a prototype rehabilitation device for hip joint that consists of three PLSAs with flexible rods and aknee supporter was proposed and tested. The universal stage for PLSA that consists of a rotational type universal
joint with radial bearings and a wire-type linear potentiometer was also proposed. For attitude control, theanalytical model that can calculate rod lengths of three PLSAs from the desired position of the knee supporter
was proposed. In addition, the method that both mechanical chucks are used as a brake in order to realize variable
PLSA was proposed. The position control system of the knee supporter using variable PLSA and the embedded
controller based on the proposed method was carried out. It could be concluded that the controlled length using
the proposed method could trace the desired length well. The standard deviation of tracking control error was
improved from 19.1 mm to 4.7 mm compared with the method using ordinal constant step motion.
To realize a compact rehabilitation device that can be used by hanging, a soft stepping actuator with bending
motion that consists of a bending unit and the improved PLSA with cross reinforced EFPAs was proposed andtested. The control system that consists of the tested PLSA, ten on/off valves and an embedded controller was also
proposed and tested. As a result, it could be confirmed that the tested PLSA can realize the rotational motion of
the knee supporter so as to trace a desired orbit, even if the more compact size of the device is used than theprevious one.
As a future work, it is necessary to investigate the suitable force to give a passive exercise for patient with various
degrees of physical disability. After that, it is needed to apply a force control to the tested rehabilitation device.

ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Numbers 16K06202 and 19K04265.

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REFERENCES
1. Annual Report on the Ageing Society [Summary] FY 2019 from, https://www8.cao.go.jp/kourei/english/annualreport/
2019/pdf/2019.pdf, June 2019.
2. NAGATA, Y., Soft Actuators, Forefront of Development. NTS Ltd., 2004, pp.291–335.
3. NORITSUGU, T., Pneumatic Soft Actuator for Human Assist Technology. In: Proceedings of the 6th JFPS International
Symposium on Fluid Power, Tsukuba, Japan, 2005, pp.11–20.
4. SASAKI, D., NORITSUGU, T., YAMAMOTO, H., TAKAIWA, M., Development of Power Assist Glove using
Pneumatic Artificial Rubber Muscle. Journal of the Robotics Society of Japan, 2006, 24(5), 640-646.
5. NORITSUGU, T., TAKAIWA, M., SASAKI, D., Development of Power Assist Wear Using Pneumatic Rubber Artificial
Muscles. Journal of Robotics and Mechatronics, 2008, 21(5), 27–31.
6. KOBAYASHI, H., SHIBAN, T., ISHIDA, Y., Realization of all 7 motions for the upper limb by a muscle suit. Journal of
Robotics and Mechatronics, 2004, 16(5), 504–512.
7. TANIGUCHI, H., WAKIMOTO, S., SUZUMORI, K., Development of a hand rehabilitation system to prevent contracture
for finger joints based on the therapy of occupational therapists (Massage a hand and range of motion exercises using
pneumatic soft actuators). Journal of the Japan Society of Mechanical Engineers, 2014, 80(820), TRANS0348.
8. KAWASAKI, H., ITO, S., ISHIGURA, Y., NISHIMOTO, Y., AOKI, T., ABE, T., SAKAEDA, H., OJIKA, T., MOURI,
T., UEKI, S., Hand Rehabilitation Assist Robot Designed with Assistance for Thumb Opposability. Journal of the Japan
Society of Mechanical Engineers, 2008, 74(748), 3019-3027.
9. TAKAIWA, M., NORITSUGU, T., SASAKI, D., Wrist Rehabilitation Using Pneumatic Parallel Manipulator ~ Proposalof Rehabilitation Based on EMG signal ~. Transactions of the Japan Fluid Power System Society, 2012, 43(3), 85-91.
10. EGUCHI, Y., AKAGI, T., DOHTA, S., “Preliminary Design of Flexible Linear Stepping Actuator Driven by Pneumatic
Balloons and Brakes”, MATEC Web of Conferences, Vol.51, No.02004, 2016, pp.1-4.
11. EGUCHI, Y., AKAGI, T., DOHTA, S., KOBAYASHI, W., “Improvement of Flexible Linear Stepping Actuator Driven
by Pneumatic Balloons and Brakes”, MATEC Web of Conferences, Vol.82, No.01005, 2016, pp.1-6.
12. FUKUKAWA, N., AKAGI, T., DOHTA, S., KOBAYASHI, W., EGUCHI, Y., “Development of Flexible Robot Arm with
Backdrivability Using Flexible Linear Stepping Actuators”, International Journal of Mechanical Engineering and Robotics
Research, Vol.6, No.5, 2017, pp.373-377.
13. FUKUKAWA, N., AKAGI, T., DOHTA, S., KOBAYASHI, W., EGUCHI, Y., “Improvement of Pneumatic Chuck in
Flexible Linear Stepping Actuator with Backdrivability”, International Journal of Mechanical Engineering and Robotics
Research, Vol.8, No.3, 2019, pp.437-442.
14. INOUE, R., AKAGI, T., DOHTA, S., KOBAYASHI, W., EGCHI, Y., FUKUKAWA, N., SHINOHARA, T.,“Development of Simple Rehabilitation Device Using Flexible Linear Stepping Actuators”, International Journal of
Mechanical Engineering and Robotics Research, Vol.8, No.2, 2019, pp.316-322.
15. MATSUI, Y., AKAGI, T., DOHTA, S., KOBAYASHI, W., TAMAKI, H., Development of Flexible Spherical Actuatorwith 3D Coordinate Measuring Device. Journal of Flow Control Measurement & Visualization, 2018, 6-2, pp.95–106.
16. SUZUKI, Y., AKAGI, T., DOHTA, S., KOBAYASHI, W., TAMAKI, H., MATSUI, Y., SHIMOOKA, S., Developmentof Tetrahedral Type Rehabilitation Device Using Flexible Pneumatic Actuators. International Journal of MechanicalEngineering and Robotics Research, 2018, 7-4, pp.409-414．
17. SHIMOOKA, S., AKAGI, T., DOHTA, S., KOBAYASHI, W., SHINOHARA, T., Improvement of Home Portable
Rehabilitation Device for Upper-Limbs. JFPS International Journal of Fluid Power System, 2019, 12-1, pp.10-18.

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Development of Extension Type Flexible Pneumatic Actuator with
Displacement Sensor Using Ring-shaped Magnet and Hall Sensor
for Tetrahedral-type Soft Mechanism
Kenshiro TAKEUCHI*, Takumi KOBAYASHI*, Tetsuya AKAGI*, Shujiro DOHTA*,
Takashi SHINOHARA**, Wataru KOBAYASHI* and So SHIMOOKA****

Department of Intelligent Mechanical Engineering, Faculty of Engineering, Okayama University of Science**
Science Dream lab., Okayama University of Science
1-1, Ridai-cho, Kita-ku,Okayama, 700-0005 Japan(E-mail: akagi@are.ous.ac. jp)***
Graduate School of Natural Science and Technology, Okayama University
3-1-1 Tsushima-naka, Kita-ku, Okayama 700-8530, Japan

Abstract. A soft mechanism that has large moving area and compliance are attractive. This study aims to develop apneumatic drive soft actuator with 3-dimensional large moving area. Based on the concept that a tetrahedron is aminimum element of solid body, the tetrahedral-type soft actuator that consists of six reinforced Extension type
Flexible Pneumatic Actuators (EFPA) and four connectors for vertices was proposed and tested. Each EFPA is set atthe position of side of tetrahedral frame. Both ends of each EFPA are connected to connectors in vertices. In addition,
to measure the displacement of the EFPA, a non-contact type displacement sensor using a ring-shaped magnet and ahall sensor was proposed. The measurement of the attitude of the tested actuator was carried out. As a result, it could
be confirmed that the tested EFPA with displacement sensor could measure whole length of the EFPA by measuring
a part of displacement of the actuator.
Keywords: Extension type flexible pneumatic actuator, Tetrahedral-type soft mechanism, Non-contact typedisplacement sensor using ring-shaped magnet and hall sensor, Measuring system using embedded controller.

INTRODUCTION
In Japan, the ratio of the elderly became more than 26 % in 2020[1]. The ratio will be expected about 36% in
2040. Based on this issue, various human friendly pneumatic devices to assist welfare work for the elderly anddisabled using pneumatic soft actuators[2] have been actively developed[3-10] As a human friendly pneumatic
drive system, K. Yamamoto developed a wearable power assist suit for nursing care using bellows type
actuators[3]. As a system using McKibben type pneumatic artificial muscle (PAM for short), T. Noritsugu and H.
Kobayashi also developed power assisted wears [4-7] to increase the force for welfare workers and heavy labors.
A rehabilitation device that can enhance to recover their physical ability early after injury and surgery is also
required. Now, Occupational Therapist (“OT” for short) and Physical Therapist (“PT” for short) can perform therehabilitation to the patient for just 30 minutes per day according to Japanese law. Some of them recommend
patients to execute voluntary rehabilitation and exercise at home. According to a medical report, voluntary
rehabilitation might prevent bedridden state, various disease and disuse syndrome [8]. Therefore, a home-based
rehabilitation device that can safely give passive exercise to patients is strongly required. As research on such arehabilitation device, a rehabilitation device for inhibiting contracture of finger joints were developed by H.
Taniguchi and H. Kawasaki [9-10]. M. Takaiwa also developed a rehabilitation device for contracture patientsusing pneumatic parallel manipulator [11]. In the previous study, the portable rehabilitation device using flexible
spherical actuator that can provide passive exercise for the wrist by changing the position of holding hands was
proposed and tested [12-14]. In these devices, a pneumatic actuator was often used because it has advantage ofcompliance based on air compressibility. A pneumatic soft actuator has more advantages that are lighter weightand more compliance based on flexibility. However, a device using soft actuators is usually difficult to realize
larger moving area because of their lower stiffness. Especially, 3-dimensional motion with flexibility is moredifficult to realize.
In this study, as a final target of the study, we aim to realize a home-based rehabilitation device that can givepassive exercise and be safely used at home with non-specialist. We assume that the home-based rehabilitation
device is used as a voluntary rehabilitation by patient, and not be used as a medical treatment by PT and OT. We

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aim to keep the moving range of the recovered joint after medical treatment by giving passive movement. In order
to use it at home for a personal user, the device must be constructed compactly and at low-cost. In the previous
study, to realize a compact home-based rehabilitation device, a compact soft mechanism that can give 3dimensional large moving area was proposed and tested. In addition, to control the attitude of the device, a soft
actuator in the soft mechanism needs a function to measure its displacement. In this paper, we propose an
extension type flexible pneumatic actuator with non-contact type displacement sensor using a ring-shaped magnet
and a hall sensor. An attachment type displacement sensor that can be applied to the already assembled soft
mechanism using EFPAs is also proposed. The measuring test using the tested mechanism with sensor is carriedout by using an unexpansive measuring system using an embedded controller.

EXTENSION-TYPE FLEXIBLE PNEUMATIC ACTUTOR
To realize a soft mechanism with a large moving area, a compact soft actuator with a long stroke motion is required.
As a pneumatic soft actuator, a McKibben type artificial muscle is well known. However, the moving area of the
McKibben type muscle is very small, the stroke is less than 25 % of its original length. Based on this issue, T.
Noritsugu developed an extension type pneumatic artificial muscle by using bellows type nylon mesh and rubber
tube [6]. The extensional performance of the muscle depends on bellows. Recently, a superior bellows type sleeve
is easily obtained as an ordinal water supply hose. In the previous study, an Extension-type Flexible Pneumatic
Actuator (we call it “EFPA” for short) using this type of hose was proposed and tested [14-16]. Figure 1 shows
the view and schematic diagram of EFPA. The EFPA consists of a silicone rubber tube with inner diameter of 8mm, outer diameter of 11 mm and covered with a bellows type nylon sleeve (The Fit Life Co. Ltd., Hose Reel).
The original length of EFPA is 200 mm, and the mass is 50 g. The actuator can be constructed at low cost, thematerial cost of the EFPA is about 5 US dollars per 1 m. The EFPA can extend about 2.5 times of its original
length. The maximum pulling force of the EFPA is about 60 N [16-17]. The pushing force of the original EFPA
is not available because the bending stiffness is too low.

Figure 1. View and schematic diagram of EFPA [14-16].

Therefore, to increase the stiffness of EFPA, a reinforced EFPA was also proposed and tested in the previous
study [18]. The left figure on Figure 2 shows the schematic diagram of the reinforced EFPA with circumferential
restraints, we call it “reinforced EFPA” for short [18]. The reinforced EFPA that consists of parallel arranged
three EFPAs as shown in the right figure on Figure 2 has restrained each other by using restraint PET plates withthickness of 1 mm at the point of sleeve. Each EFPA keeps so as to arrange every 120 deg. at radius of 12.5 mm
from the center of the reinforced EFPA. By this method, the bending stiffness of the actuator increases, and thereinforced EFPA can generate the pushing force of 130 N [18].

Figure 2. Schematic diagram of reinforced EFPA on the left and Shape of reinforced PET plate on the right.

TETRAHEDRAL-TYPE SOFT MECHANISM
By paying attention to the fact that a tetrahedron is a minimum element of a solid body, a tetrahedral-shaped soft
actuator using EFPAs that can form an arbitrary shape by extending each side of the tetrahedron was also proposed2

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and tested. Figure 3 shows a view of a tested tetrahedral-type soft mechanism. The tested mechanism consists ofsix reinforced EFPAs with a natural length of 130 mm and four connectors for vertex. The vertex is made of ABS
resin. Each reinforced EFPA is set at the position of side of the tetrahedral frame. Both ends of each reinforced
EFPA are connected to connectors in vertices. Three EFPAs in the reinforced EFPA are connected in series by asupply tube and operated simultaneously. Since the reinforced EFPA has flexibility, it can be naturally curved
even if the rigid connector for vertex is used. By this method, the shape of the tested mechanism can change tovarious shape of tetrahedron.

Figure 3. Tetrahedral-type soft mechanism using reinforced EFPAs.

Figure 4 shows the configuration of the sequential control system of the tested mechanism. The system consistsof the tested mechanism, six sets of operating valves that consists of two on/off valves (Koganei Co. Ltd., G010E1), and an embedded controller (Renesas Electronics Co. Ltd., SH7125). In each operating valve, the output portof the switching three port on/off valve is connected to supply port of two-port on/off valve for holding. Eachon/off valve is driven through I/O ports in the embedded controller and transistors and can perform threeoperations, that is to “supply”, “exhaust” and “hold”.

Figure 4. Control system of the tetrahedral-type soft mechanism.

Figure 5 shows the view of movement of the tested mechanism when the sequential operation was applied. As
shown in Figure 5 (2), when the three reinforced EFPAs connected with top vertex connector was pressurized, themechanism extends until 350 mm in height. In opposite case when the three reinforced EFPAs on the bottom werepressurized as shown in Figure 5 (3), the mechanism contracts toward vertical direction. In this case, the minimum
height of the mechanism of 195 mm can be realized. In addition, as shown in Figure 5 (5), the mechanism canbend toward six radial directions by pressurizing one or two reinforced EFPAs connected with the top vertex
connector. As an application using the mechanism, we think that it can be used as a portable rehabilitation device.
By a patient holding the two vertexes connector while the mechanism deforms to various shapes, the actuator cangive passive exercise to patient’s wrist and upper limbs. Furthermore, by combining more tetrahedral-type soft
mechanisms to construct polyhedron, a polyhedral mechanism can give complex passive exercise to a patient.

Figure 5. View of movement of the tested actuator when sequential operation was applied.

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MEASUREMENT OF MECHANISM
To measure the shape of the mechanism, several trials for measurement were carried out. In one of them, by
paying attention that the shape of tetrahedron can be roughly imaged by the position of top vertex, a measuring
device to recognize the position of top vertex is proposed and tested. The left figure on Figure 6 shows theconstruction of the tested measuring device. The device consists of a rotational type two axis universal joint withtwo potentiometers and wire-type potentiometer [14]. The end of wire of three potentiometers are connected tothe behind of the top vertex connector. The operating principle is as follows. In ideal case of no friction, under thecondition when the top vertex connector is always pulled by the linear potentiometer, the universal stage is always
directed toward the top vertex. In the universal stage, the bottom side potentiometer can measure the directional
angle, that is bending directional angle of the mechanism. The side potentiometer can measure an elevation angle,
that is bending angle of the mechanism according to the position change of the top vertex. In addition, the lengthfrom the measurement origin can be detected by the wire-type linear potentiometer. By obtaining two kinds ofbending angle and a distance, the three-dimensional coordinate can be recognized. The right schematic diagramon Figure 6 shows the measuring system that was reconstructed on the previous sequential control system asshown in Figure 4.

Figure 6. View and system of measurement device using rotational universal stage and wire type linear potentiometer.

Figure 7 (a) and (b) show the transient response of bending angles and the length from the measurement origin,respectively. In Figure 7 (a), red and blue lines show results of the bending directional angle and the bending
angle, respectively. It can be seen that the tested measuring device can detect enough information to calculate thecoordinate based on polar coordinate system. However, because the real tested device has friction, the measuring
error is occurred. Especially the measuring error of bending directional angle is easily occurred because the inertia
of universal stage is large. In addition, because the liner potentiometer always is pulling the top vertex connector,
the deformation of the mechanism is always affected by the measuring device.

(a) Bending directional angle and bending angle.

(b) Length from measuring origin.

Figure 7. Transient response of bending directional angle, bending angle and length from measuring origin.

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NON-CONTACT TYPE DISPLACEMENT SENSOR
Therefore, a non-contact type displacement or position sensor is required. As a measuring method, the visual
camera is one of available methods. However, the camera detecting system depends on light condition and camera
position. In addition, such a measuring system using camera requires a superior personal computer for calculationand it is very expensive. It is also very difficult to realize a mounted measuring system using camera. In this study,
we tried to develop of a non-contact type displacement sensor of each EFPA. As a specification of the sensor,
because EFPA is so flexible, the required sensor needs an ability to measure the displacement of EFPA even if
EFPA bends and twists. Therefore, a novel non-contact type displacement sensor that can measure a longitudinal
displacement under the condition when the radial directional deviation is occurred is proposed. The left figure on
Figure 8 shows the fundamental construction of the proposed sensor. The sensor consists of a hall sensor (Allegro
Co. Ltd., A1324LUA-T) and a ring-shaped neodymium magnet with the outer diameter of 15 mm, the innerdiameter of 9 mm and the height of 1 mm. The hall sensor is set on the longitudinal axis of the ring-shaped magnet.
By using the ring-shaped magnet, the stable cylindrical magnetic field not related to radial directional positionchange can be used. The right graph on Figure 8 shows the relation between the longitudinal position(displacement) and the output from the hall sensor. The displacement of 0 mm shows the position where the S
pole side surface of the ring-shaped magnet meets the surface of the hall sensor. In this figure, each symbol shows
the radial directional deviation from the center axis. From Figure 8, it can be seen that the small radial deviationof less than 2 mm is not affected to measure longitudinal displacement. It means that the tested sensor can measure
the longitudinal displacement even if EFPA bends or twists. However, it can also be seen that the measuring areaof hall sensor is not so large, that is less than 7 mm. Therefore, we apply the tested sensor to measure thedisplacement of one bellows in EFPA.

Figure 8. Fundamental construction and static characteristics of non-contact type displacement sensor using ring-shaped
magnet and hall sensor.

In the same manner of the previous experiment, the relations between the longitudinal position (displacement)
and the output from the hall sensor using various ring-shaped magnets were investigated. Table 1 shows the result,that is the measuring area of the hall sensor using various ring-shaped magnets. In the experiment, 16 types ofring-shaped magnets were used. In order to measure the displacement of one bellows in EFPA, it is necessary
more than 7 mm as a measuring area. In addition, it is better to use small-sized magnet, because it must be attachedon the EFPA. From view point of wider measuring area and small size magnet, we select the magnet with theouter diameter of 15 mm, the inner diameter of 9 mm and width of 1 mm.
Table 1. Measuring area of hall sensor for various ring-shaped magnets.

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EFPA WITH NON-CONTACT TYPE DISPLACEMENT SENSOR
Figure 9 shows the construction and the schematic diagram of a reinforced EFPA with the non-contact typedisplacement sensor. The hall sensor is set at the position with radius of 24 mm from the center of the restraint
plate in the reinforced EFPA. The ring-shaped magnet is also set on anther restraint plate at the same position.
Both restraint plates for the ring-shaped magnet and the hall sensor are set on reinforced EFPA through one
bellows. As assuming that the behavior of each bellows is same at any points, the sensor can estimate the wholedisplacement by measuring the displacement of one bellows.

Figure 9. Construction and schematic diagram of reinforced EFPA with non-contact type displacement sensor.

In order to calculate the displacement from sensor output, the calibration test of the tested sensor was carried out.
Figure 10 (a) shows the relation between the hall sensor output and the whole displacement of the reinforced
EFPA. In the experiment, so as to get increasing output value according to increasing displacement, the pole of
the ring-shaped magnet was inversed compared with the previous case as shown in Figure 8. The output voltage
from the hall sensor was detected through A/D converter in the embedded controller (Renesas electronics Co. Ltd.,
SH7125). Based on this relationship, the following approximate equation can be obtained.
𝑥 = 8.91 × 10−7 𝑠 3 − 7.75 × 10−4 𝑠 2 + 0.326𝑠 − 9.63

(1)

where s and x mean the A/D value of hall sensor output voltage through 10-bit A/D convertor in the embedded
controller and the whole displacement of the reinforced EFPA, respectively. This cubic equation is selected as
approximate equation in order to be easily calculated by a tiny embedded controller.
Figure 10 (b) shows the transient response of the displacement of the reinforced EFPA when the supply pressureis changed from 0 to 500 to 0 kPa manually. In Figure 10, the blue solid line and red broken line show the true
displacement that is measure by the connected linear potentiometer (Copal Co. Ltd, JCL100B) in serial and themeasure displacement using the sensor output and the approximate model, respectively. From Figure 10, it can beseen that the measured displacement using the tested sensor is relatively good agreement with the true
displacement although the small error exists while exhausting. The standard deviation of measuring error is about
4.5 mm, that is about 5 % error of full range. It seems that the error is occurred by uneven contracting of the rubber
tube under the condition of large frictional force based on higher pressure. As a result, it can be confirmed thatthe proposed method to measure whole displacement of EFPA by measuring a part of EFPA using the tested
sensor is valid.

(a) Result of calibration test using the tested sensor.

(b) measuring result using tested sensor.

Figure 10. Relation between sensor output and whole displacement of the reinforced EFPA and transient response of thedisplacement of the EFPA when the supply pressure is changed from 0 to 500 to 0 kPa.

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The previous measuring method mentioned above requires the calibration from the sensor output to the wholedisplacement of the EFPA, each time EFPA changed. In order to improve versatility of displacement measurement,
the method to estimate the whole displacement from the measured displacement of one bellows by using the tested
sensor is carried out. From the fundamental relationship between sensor output and displacement as shown in
Figure 11(a), the following approximate equation can be obtained.

𝑥 = 𝑛 × {4.70 × 10−8 𝑠 3 − 3.84 × 10−5 𝑠 2 + 1.56 × 10−2 𝑠 − 0.517}(2)

where s, x and n mean the sensor output (A/D value) and the whole displacement of EFPA and number of bellows
in the EFPA, respectively.Figure 11 (b) shows the relation between the supply pressure and the calculated andtrue displacement of the reinforced EFPA. In Figure 10, symbols and broken line show the calculated and true
whole displacement of the reinforced EFPA, respectively. The calculated displacement is based on the sensor
output and the approximate equation mentioned above. The true displacement is the measured displacement byusing linear potentiometer that connected to the reinforced EFPA in serial. Symbols ● and▲ show the case when
number of bellows in EFPA is estimated as 21 and 20, respectively. In the experiment, although the number ofbellows is 20, the number of 21 was used as a calculated value in consideration of looseness of bellows and rubber
tube length. From Figure 11 (b), it can be seen that the calculated displacement almost meets the true displacement
even if the relation between the supply pressure and displacement has a large hysteresis. It can also be seen that
there is relative larger measuring error while exhausting because of same reason mentioned above. Overall, it canbe seen that the calculated displacement can predict the true displacement relative well.

(a) Result of calibration test using the tested sensor.

(b) Measured result using tested sensor.

Figure 11. Relation between the supply pressure and displacement of the reinforced EFPA.

CONCLUSIONSAs an attitude sensor for the tetrahedral shaped soft mechanism, the non-contact type displacement sensor using
a hall sensor and a ring-shaped magnet was proposed and tested. As a result of fundamental experiment, it foundthat the ring-shaped magnet with outer diameter of 15 mm and the inner diameter of 9.1 mm and width of 1 mm
is suitable to use as a displacement sensor of EFPA from view point of small size and larger measuring area. It
also found that the tested sensor can measure the longitudinal displacement even if a radial deviation of less than
2 mm is occurred by using the stable cylindrical magnetic field. The measuring system using the tested sensor andthe embedded controller was proposed and tested. The reinforced EFPA with the tested sensor for measuring thedisplacement of one bellows was proposed and tested. The measuring methods to measure whole displacement of
EFPA by measuring a part of EFPA using the tested sensor were proposed. The measuring system using the tested
sensor and the embedded controller was also developed. As a result, by using the method to use the approximate
equation from sensor output to the whole displacement, it can be confirmed that the measured displacement using
the tested sensor was relatively good agreement with the true displacement with the standard deviation of
measuring error of about 4.5mm, that is corresponding to about 5 % error of full range. As a more versatilemeasuring method, the method to estimate the whole displacement from the measured displacement of one
bellows was also carried out. As a result, it can also be confirmed that the calculated displacement almost meetsthe true displacement even if the relation between the supply pressure and displacement has a large hysteresis.
In the future work, we are going to develop an unite type displacement sensor for EFPA. The attitude control of
the tetrahedral shaped soft mechanism using the improved EFPAs with sensor was carried out. In addition, it isnecessary to investigate the measuring error under the condition of load.

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ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Numbers 16K06202 and 19K04265.

REFERENCES
1. Annual Report on the Ageing Society [Summary] FY 2019 from, https://www8.cao.go.jp/kourei/english/annualreport/
2019/pdf/2019.pdf, June 2019.
2. NAGATA Y. Soft Actuators, Forefront of Development. NTS Ltd., 2004, pp 291–335.
3. ISHII M, YAMAMOTO K and HYODO K. Stand-Alone Wearable Power Assist Suit (Development and Avail ability).
Journal of the Japan Society of Mechanical Engineers, 2006, 72-715, pp175–182.
4. NORITSUGU T. Pneumatic Soft Actuator for Human Assist Technology. In: Proceedings of the 6th JFPS International
Symposium on Fluid Power, Tsukuba, Japan, 2005, pp11–20,
5. SASAKI D, NORITSUGU T, YAMAMOTO H and TAKAIWA M. Development of Power Assist Glove using Pneumatic
Artificial Rubber Muscle. Journal of the Robotics Society of Japan, 2006, 24-5, pp640-646.
6. NORITSUGU T, TAKAIWA M and SASAKI D. Development of Power Assist Wear Using Pneumatic Rubber Artificial
Muscles. Journal of Robotics and Mechatronics, 2008, 21-5, pp27–31.
7. KOBAYASHI H, SHIBAN T and ISHIDA Y. Realization of all 7 motions for the upper limb by a muscle suit. Journal of
Robotics and Mechatronics, 2004, 16-5, pp504–512.
8. Council of Bureau of Social Welfare and Public Health Tokyo Metropolitan Government from, https://www.
fukushihoken.metro.tokyo.lg.jp/iryo/sonota/riha_iryo/kyougi01/rehabiri24.files/siryou242.pdf; February 2013 (inJapanese).
9. TANIGUCHI H, WAKIMOTO S, SUZUMORI K. Development of a hand rehabilitation system to prevent contracture forfinger joints based on the therapy of occupational therapists (Massage a hand and range of motion exercises using
pneumatic soft actuators). Journal of the Japan Society of Mechanical Engineers, 2014, 80-820, TRANS0348.
10. KAWASAKI H, Ito S, ISHIGURA Y, NISHIMOTO Y, AOKI T, ABE T, SAKAEDA H, OJIKA T, MOURI T, UEKI S.
Hand Rehabilitation Assist Robot Designed with Assistance for Thumb Opposability. Journal of the Japan Society of
Mechanical Engineers, 2008, 74-748, pp3019-3027.
11. TAKAIWA M, NORITSUGU T, SASAKI D. Wrist Rehabilitation Using Pneumatic Parallel Manipulator ~ Proposal of
Rehabilitation Based on EMG signal ~. Transactions of the Japan Fluid Power System Society, 2012, 43-3, pp85-91.
12. MATSUI Y, AKAGI T, DOHTA S, ALIFF M, LIU C. Development of Portable Rehabilitation Device Using Flexible
Spherical Actuator and Embedded Controller. Lecture Notes in Electrical Engineering, 2014, 293-1, pp231-238.
13. MATSUI Y, AKAGI T, DOHTA S. Development and Control of Flexible Spherical Actuator for Portable Rehabilitation
Device. International Journal of Materials Science and Engineering, 2015, 3-1, pp49-54.
14. MATSUI Y, AKAGI T, DOHTA S, KOBAYASHI W, TAMAKI H. Development of Flexible Spherical Actuator with 3D
Coordinate Measuring Device. Journal of Flow Control Measurement & Visualization, 2018, 6-2, pp95–106.
15. SHIMOOKA S, DOHTA S, AKAGI T, KOBAYASHI W and YONEDA M. Development of Portable Rehabilitation
Device Using Flexible Extension Type Soft Actuator with Built-in Small-Sized Quasi Servo Valve and Displacement
Sensor. JFPS International Journal of Fluid Power System, 2018, 11-2, pp18-25.
16. SUZUKI Y, AKAGI T, DOHTA S, KOBAYASHI W, TAMAKI H, MATSUI Y, and SHIMOOKA S. Development of
Tetrahedral Type Rehabilitation Device Using Flexible Pneumatic Actuators. International Journal of MechanicalEngineering and Robotics Research, 2018, 7-4, pp409-414．
17. SHIMOOKA S, DOHTA S, AKAGI T, KOBAYASHI W and SHINOHARA T. Improvement of Home Portable
Rehabilitation Device for Upper-Limbs. JFPS International Journal of Fluid Power System，2019, 12-1, pp10-18.
18. SHIMOOKA S, DOHTA S, AKAGI T, SHINOHARA T. HANE Y and ALIF M. Development of Reinforced Extension
Type Flexible Pneumatic Actuator with Circumferential Restraints and Its Application for Rehabilitation Device.
International Journal of Automotive and Mechanical Engineering, 2020, 17-3, pp8116-8127.

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Development of Highly Durable Straight Fiber Type Pneumatic
Artificial Muscle with a Double Structural Air Chamber
Naoki SAITO*, Daisuke FURUKAWA*, Toshiyuki SATOH* and Norihiko SAGA***

Department of Intelligent Mechatronics, Akita Prefectural University
84-4, Ebinokuchi, Tsuchiya, Yurihonjo, Akita, 015-0055, Japan(E-mail: naoki_saito@akita-pu.ac.jp)**
Department of Human System Interaction, Kwansei Gakuin University

Abstract. In this paper, we describe a straight fiber type pneumatic artificial muscle with high durability. The airchamber has a double structure of a high strength airbag which has been investigated in Rubber Less Artificial Muscle
and rubber tube. The relationship between the internal pressure and contraction displacement, and inner pressure andcontraction force was experimentally investigated, and it was confirmed that the hysteresis was less than that of theconventional Rubber Less Artificial Muscle. In addition, position control by this proposed artificial muscle was
performed and good characteristics was obtained.
Keywords: Pneumatic, Actuator, Pneumatic Artificial muscle, Basic characteristics, Position control

INTRODUCTION
The development and application of pneumatic artificial muscles, one of the soft actuators, have been promotedin the new mechatronics field such as soft robotics [1-4]. While general pneumatic rubber artificial muscles arelightweight, and it can generate a large contraction force. However, it is considered that the practicality as
actuators is poor because of their short durability due to softness of the rubber material.
On the other hand, the Rubber Less Artificial Muscle (RLAM) proposed by Saito et al. [5] has shown durabilityof more than 200,000 cycles under practical loads and driving pressures [6], indicating its potential as a practicalactuator.
However, since RLAM only actuate in the direction of contraction, constraints such as the antagonistic placementof two artificial muscles are necessary [6], it is difficult to use for control involving extension and contraction.
In addition, in the extension and contraction behavior against a load, it contracts when the inner pressure increases,
but it does not extend easily when the pressure is decreased. Therefore, it is known to have a hysteresis
characteristic [7]. This characteristic makes the design of the control system very difficult, so it is desirable toeliminate the hysteresis characteristic.
Therefore, in this study, we propose a pneumatic artificial muscle with a double structural air chamber consistingof a high strength airbag, which has been developed in RLAM, and a rubber tube, which is expected to reduce
hysteresis caused by extension movements due to elastic deformation. This artificial muscle will be fabricated,
and the reduction of hysteresis and control performance will be verified to confirm the usefulness of the actuator.

STRUCTURE OF PROPOSED ARTIFICIAL MUSCLE
In this study, we fabricate a pneumatic artificial muscle called the straight fiber type [8, 9], as shown in Figure 1.
This pneumatic artificial muscle has a relatively large contraction ratio. In addition, this type has a structuraldisadvantage in that the deformation of the rubber is larger and more prone to tearing or rupture because of thefibers used to constrain the deformation of the pneumatic artificial muscle are arranged along the axial direction.
As shown in Fig. 2, the inside of the air chamber is composed of a high strength airbag, and the outside is covered
with a rubber tube to form a double structured air chamber. The axial fibers were used of Kevlar fiber, and eightfibers were arranged so that they were evenly distributed in the circumferential direction as shown in Fig. 2 andattached to the inner airbag so that they slid in the axial direction. A ring made of Kevlar fiber equal to thecircumferential length was attached at two axial locations. The amount of axial contraction caused bycircumferential expansion increases and enables deformation with a higher contraction ratio. In this way, theHighly Durable Straight Fiber type pneumatic Artificial Muscle (HD–SFAM) was fabricated. The actual
fabricated HD-SFAM is shown in Figure 3, and the main parameters are listed in Table 1. By applying a high1

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Figure 1. Basic mechanism and drive principle of the straight fiber type pneumatic artificial muscle

Figure 2. Detailed structural diagram of the HD-SFAM

contraction

extension
Figure 3. Prototype HD-SFAM

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Table 1. Parameters of prototype HD-SFAM
23.5
22.5
57.3
210.0

Outside diameter of Rubber tube
Inside diameter of Rubber tube
Diameter of High strength air bag
Lengthstrength airbag on the inner side, the HD–SFAM is considered to be able to stretch and contract about 200,000
times [6], basically the same as RLAM.
In the next chapter, we will examine the basic characteristics of this HD–SFAM.

BASIC CHARACTERISTICS
Relationship between inner pressure and contraction ratio
In this section, we first investigate the relationship between inner pressure and contraction ratio as a basic
characteristic. The experimental system is shown in Fig. 4. The inner pressure is applied to HD–SFAM with theload suspended, and the contraction ratio is measured. The inner pressure is increased from 0 MPa to 0.2 MPa by
0.02 MPa every 5 seconds, and then decrease to 0 MPa under the same conditions. The average contraction ratio
at constant pressure is shown in Fig. 5. From the results, it can be seen that the hysteresis loop becomes smaller
as the load increases. Especially at 200 N, the hysteresis loop almost disappears.
Figure 6 shows the comparison of the pressure contraction characteristics between the conventional RLAM andthe proposed HD-SFAM when the load is 200N. This result shows that the rubber tube is effective in reducing thehysteresis. On the other hand, in Fig. 5, when the load is 50 N, the hysteresis loop is large, and the HD-SFAM
does not sufficiently reduce the hysteresis characteristics depending on the conditions. This may be due to theinsufficient contraction force of the rubber tube. The thickness of the rubber tube used in this study is 0.5 mm,
and we believe that using thicker rubber tube would increase the elastic force for stretching HD–SFAM and
improve the characteristics.

Relationship between inner pressure and contraction force(Isometric contraction characteristic)
Next, we investigate the relationship between inner pressure and contraction force when the contraction ratio was
kept constant. The experimental system is shown in Fig. 7. The experimental conditions are as follows: theFigure 4. Experimental setup for measuring the relationship between inner pressure and contraction ratio
25

Contraction ratio [%]

20
15
10

50N
100N
150N
200N

5
0
0

0.05

0.1
Pressure [MPa]

0.15

0.2

Figure 5. relationship between inner pressure and contraction ratio

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Contraction ratio[%]

20
15
10
5

RLAM 200N
HD-SFAM 200N

0
0

0.05

0.1
Pressure[MPa]

0.15

0.2

Figure 6. Comparison result of inner pressure and contraction ratio of RLAM and HD-SFAM

Figure 7. Experimental setup for measuring the relationship between inner pressure and contraction force

600
Contraction ratio 0％
Contraction ratio 5％
Contraction ratio 10％
Contraction ratio 15％
Contraction ratio 20％

500

Force[N]

400
300
200
100
0
0

0.05

0.1

0.15

0.2

Pressure[MPa]Figure 8. relationship between inner pressure and contraction force

contraction ratio is set from 0% to 20% at 5% intervals, and the inner pressure is applied from 0 MPa to 0.2 MPa
at 0.02 MPa intervals. The results are shown in Fig. 8. From the results, it can be seen that the contraction force
decreases with increasing contraction ratio. This trend is similar to that of other pneumatic artificial muscles. In
the natural length, a force of about 540 N can be output at a pressure of only 0.2 MPa. In addition, this highstrength airbag is capable of applying a pressure of 0.6 MPa [6], which means that a large maximum contraction
force can be expected by carefully considering the strength of the components such as the hose band.

POSITION CONTROLIn this section, we evaluate the control performance of the developed HD-SFAM. Considering the HD-SFAM asa mass-spring-damper system, the equation of motion becomes as follows:𝑑! 𝑥(𝑡)
𝑑𝑥(𝑡)
(1)
𝑚
+𝑐
+ 𝑘𝑥(𝑡) = 𝐹(𝑡)!
𝑑𝑡𝑑𝑡
where 𝑥(𝑡) is the amount of contraction, 𝑚 is the mass at which the force acts, 𝑘 is the elastic coefficient that
describes the spring characteristics of the rubber tube, and 𝑐 is the damping coefficient. The relationship between
the force 𝐹(𝑡) and the pressure 𝑃(𝑡) can be expressed by the following equation, using the coefficient 𝐴:
𝐹(𝑡) = 𝑃(𝑡)𝐴(2)
From Eq. (1) and Eq. (2), the following equation is obtained:

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𝐴
(3)
𝑃(𝑠) ,
𝑚𝑠 ! + 𝑐𝑠 + 𝑘
This result suggests that the proposed artificial muscle is a typical second-order delay system. Therefore, forposition control, it is necessary to incorporate an integrator in the feedback controller to eliminate the steady-state
position error, in this paper, position control is basically performed by using PI controller.
The experimental system is the same as in Fig. 4, but with two different suspending loads. One is a suspending
load of 200 N, which minimizes the effect of hysteresis. The other is a suspending load of 100 N, and the effectof hysteresis remains a little. The PI controller gain is adjusted by trial and error considering the amount ofovershoot and the settling time, and the same gain is used in the two experiments. Figure 9 shows the step responseof the position control by the PI controller with a suspending load of 100 N. And Figure 10 shows the step responseof the position control by the PI controller with a suspending load of 200 N. Both results show that no steady-state
error and no overshoot, and the settling time is about 3 s. The settling time is slightly shorter when the suspending
load is 100 N. This result indicates that since the controller gain is the same, the smaller load has less work tointerfere with the motion and thus reaches the target position faster. Figure 11 shows the result of changing thesuspension load during position control. This result shows that the position control is properly performed evenwhen the external load is changed, and the amount of contraction converges to the target value.
These results confirm that the HD-SFAM has almost no hysteresis effect and is useful as an actuator for systems
such as robot joints that require continuous position control.
𝑋(𝑠) =

Figure 9. Step response of the position feedback control with PI controller and suspending load of 100 N

Figure 10. Step response of the position feedback control with PD controller and suspending load of 200 N

Figure 11. Step response of the position feedback control with PI controller when suspending load changes

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CONCLUSION
In this study, we proposed and fabricated a highly durable straight-fiber type pneumatic artificial muscle (HDSFAM) with a double structure of high-strength airbag and rubber tube. The results obtained are summarized asfollows:
1. From the relationship between the inner pressure and the contraction ratio, it was found that the hysteresis
decreased as the load increased and almost disappeared when the load is about 200N. The hysteresis tended tobe smaller than that of rubberless artificial muscle which we proposed in previous research. However, hysteresis
still occurs depending on the load conditions. In this prototype, we were not able to eliminate the hysteresis.
2. The isometric contraction characteristics were similar to those of general pneumatic artificial muscles, and thecontraction force decreases with increasing in the contraction ratio.
3. From the results of position control, good position control results were obtained with PI controller. The control
system was hardly affected by changes in the magnitude of hysteresis due to changes in the external load, andthe response converged quickly to the target value even when the external load was changed during position
control.
These results suggest that the proposed HD-SFAM is useful as an actuator for soft robotics.
Further reduction of the hysteresis by re-examining the parameters such as the thickness of the rubber tube andthe material of the high strength airbag of the HD-SFAM is our future work.

REFERENCES
1. TAKAOKA, M., SUZUMORI, K., WAKIMOTO, S., IIJIMA, S. and TOKUMIYA, T., Fabrication of thin McKibben
artificial muscle with various design parameters and their experimental evaluations, Proceedings of 5th International
Conference on Manufacturing, Machine Design and Tribology (ICMDT2013), 2013, p.82.
2. YANAGIDA, T., ADACHI, K. and NAKAMURA, T., Development of Bellows-Type Artificial Rubber Muscle andApplication to Peristaltic Crawling Endoscopic Robot, Journal of Robotics and Mechatronics, 2013, 25-4, pp. 748-754.
3. LI, W., NABAE, H., ENDO, G. and SUZUMORI, K., New Soft Robot Hand Configuration with Combined Biotensegrity
and Thin Artificial Muscle, IEEE Robotics and Automation Letters (RAL), 2020, 5-3, pp. 4345-4351.
4. KADOWAKI, Y., NORITSUGU, T., TAKAIWA, M., SASAKI, D. and KATO, M., Development of Soft Power-Assist
Glove and Control Based on Human Intent,” Journal of Robotics and Mechatronics, 2011, 23-2, pp. 281-291.
5. SATO, T., SAITO, N., OGASAWARA, T. and SATO, T., Development of Rubberless artificial muscle and fundamental
characteristics, Proceedings of the 37th Annual Conference of the IEEE Industrial Electronics Society (IECON2011), 2011,
pp. 2051-2056.
6. SAITO, N. and SATOH, T., Structure of Rubberless Artificial Muscle and Evaluation of a Lifetime, Proceedings of the42th Annual Conference of the IEEE Industrial Electronics Society (IECON2016), 2016, pp.648-653.
7. SAITO, N., SATO, T., OGASAWARA, T., TAKAHASHI, R. and SATO, T., Mechanical equilibrium model of rubberless
artificial muscle and application to position control of antagonistic drive system Industrial Robot: An International Journal,
2013, 40-4, pp. 347-354.

8. TOMORI, H. and NAKAMURA, T., Theoretical Comparison of McKibben-Type Artificial Muscle and Novel
Straight-Fiber-Type Artificial Muscle, International Journal of Automation Technology, 2011, 5-4, pp. 544550.
9. FRANCESCO, D., MICHELE, G. A., PIERLUIGI, B. Z. and TERENZIANO, R., Development of a Straight
Fibers Pneumatic Muscle, International Journal of Automation Technology, 2018, 12-3, pp. 413-423.

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The 11th JFPS International Symposium on Fluid Power 2020

Evaluation of lifting motion with non-wearing typepneumatic power assist device~ comparison of active and passive type ~Masashi YOKOTA*, Reito HIRABAYASHI*, and Masahiro TAKAIWA*** Graduate School Advanced Technology and Science, Tokushima University** Graduate School of Technology, Industrial and Social Sciences, Tokushima University
2 1 Minami Johsanjimacho, Tokushima, 770 8506 Japan(E-mail: c501942002@tokushima-u.ac.jp)

Abstract. In Japan, we are facing an aging society. Securing the young labor force is a serious problem especially inthe nursing field to support care recipients and in an industrial field to support heavy labor. In the lifting motion, there
are roughly 2 types that are the squat method and the stoop one. The squat method is recommended because the burden
on the waist is lowered than the stoop one. Many types of power assist devices are developed at research institutes
around the world. However, most of them are wearable types and they promote the stoop method from their mechanicalproperty. In a previous study, we developed a non-wearing type pneumatic power assist device that takes the squat
method into consideration and evaluated a lifting motion with a position control system. In this study, we developed apassive assist device as well as an active one and evaluate the effectiveness in lifting motion between them.
Keywords: Pneumatic actuator, Power assist device, Force control, Disturbance observer, Human support.

INTRODUCTION
In Japan, we are facing an aging society, where the shortage of young labor causes a serious problem, especiallyin the nursing field and primary industrial one. In both working fields, we focus on a lifting motion. A back painis the most common occupational disease by research of the Japan Ministry of Health, Labor and Welfare(hereinafter referred to as “MHLW”) [1]. There are roughly 2 types in the lift motion, a squat method and a stoop
one. The former involves lifting with a bent knee and the latter involves only the waist, as shown in Figure 1.
When humans lift the object more than 15% of their body weight, they unconsciously shift from the squat method
to the stoop one [2, 3]. MHLW recommends the squat method because the burden on the waist is lower than thestoop one [4]. The power assist device is one of the solutions to cope with this problem, and many types of devices
have been developed at research institutes around the world. A wearable power assist device, which is the mosttypes of devices, decreases the burden on the waist [5~7]. But it is not easy to bend the knee since it pushes thefront part of the thigh backward to generate extension torque around the waist joint. Therefore, it is not too muchto say, a wearable power assist device may promote the stoop method. Moreover, the wearer is burdened by the
total weight of the device. In the meanwhile, an exoskeleton type device unloads the weight of the device to theground [8~10]. But it takes time to wear the device and become large scale as well as introduction cost.
In the previous study, we have developed a non-wearing type pneumatic power assist device that takes the squat
method into consideration and evaluated a lifting motion with a position control system [11]. In this study, wedeveloped a passive type assist device as well as active one. Since the energy consumed by an active type device
is greater than that of a passive one, it's not easy to simply compare them. However, we evaluate the effectivenessin both types of the devices from view of reducing physical burden. After presenting the schematics of both typesof assist devices, the effectiveness in lifting motion was evaluated through some experiments.

(a) Squat method(b) Stoop method
Figure 1. Lifting motion type.

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NON-WEARING TYPE PNEUMATIC POWER ASSIST DEVICE
Overview of the developed device
The 2 types of developed power assist device (hereinafter referred to as "the assist device") is shown in Figure 2
and 3. As you see that, both assist devices contact with wearer at just 2 points. One is shoe side and contact withaground and the other pushes up wearer’s armpit. Consequently, the wearer does not have physical constraints
with the assist device and can lift an object by the squat method. The average height of Japanese persons in their
20s to 50s is approximately 1.65 m [12] and the armpit is located at approximately 82% of the height [13].
Therefore, the supporting position (armpit) of the assist device is approximately 1.35 m. Since a parallel linkmechanism is introduced as shown in Figure 4, the assist device behaves as one D.O.F link mechanism.
The active type power assist device shown in Figure 2 is using wire-type pneumatic cylinders. The maximum
rotational torque of the actuator is 55.8 Nm at a cylinder pressure of 450 kPa. The total weight of the device isapproximately 5.0 kg. Figure 5 shows the pneumatic driving circuit for active type. The rotational angle of thepulley 𝜃 and pressure in the cylinder chambers (𝑝1 , 𝑝2 ) are detected by a potentiometer and pressure sensors(Keyence Corp., AP-C43), respectively. The output voltages from these sensors are fed back to an A/D converter.
A flow type servo valve (FESTO, KMPYE-5) regulates the pressures 𝑝1 and 𝑝2 in the cylinder. The control
system is implemented using RTAI, a real-time extension of Linux, and the sampling period is set to 5 ms.
The passive type power assist device shown in Figure 3 has just a spring instead of an actuator. The support force
depends on the number of a spring. The total weight of the device is approximately quite light of 1.8 kg. Since thedevice consist of only passive element of springs, a subject has a burden when they are crouching.

(a) Appearance of motion(b) Appearance of device(a) Appearance of motion(b) Appearance of device

Figure 2. Active type power assist device

Figure 3. Passive type power assist device

Figure 4. Kinematic model

Figure 5. The pneumatic driving circuit for active type2

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Kinematic analysis
Figure 4 shows the kinematic model of the assist device where a base coordinate system 𝑂 is set at the lower end
and hand vector (the relative position of armpit to the base coordinate system) and joint vector is defined asℎ = [𝑥, 𝑦]𝑇 and 𝜃 = [𝜃1 , 𝜃2 , 𝜃3 ]𝑇 , respectively. Forward kinematics ℎ = 𝐼(𝜃) is described as follows.
{

𝑥 = 𝑎 cos(𝜃1 + 𝜃3 − 𝜃2 ) − 𝑏 cos(𝜃2 − 𝜃3 ) + 𝑐 cos(𝜃3 )𝑦 = 𝑎 sin(𝜃1 + 𝜃3 − 𝜃2 ) + 𝑏 sin(𝜃2 − 𝜃3 ) + 𝑐 sin(𝜃3 )(1)

Due to the introduced parallel link mechanism, a joint angle is constrained as 𝜃1 = 𝜃2 . Consequently, the forward
kinematics is rewritten as follows.
{

𝑥 = 𝑎 cos(𝜃3 ) − 𝑏 cos(𝜃1 − 𝜃3 ) + 𝑐 cos(𝜃3 )𝑦 = 𝑎 sin(𝜃3 ) + 𝑏 sin(𝜃1 − 𝜃3 ) + 𝑐 sin(𝜃3 )(2)

The relation between a torque vector 𝜏 = [𝜏1 , 𝜏2 ]𝑇 a torque around each joint and an applying force vector at the
supported position (contact point with armpit) 𝐹 = [𝐹1 , 𝐹2 ]𝑇 is written in the following equation using a Jacobianmatrix 𝐽𝑎𝑐𝑜 from a principle of virtual work.
𝜏 = 𝐽𝑎𝑐𝑜 𝑇 𝐹(3)Expanding Eq.3 gives the following statics balance equation.
{

𝜏1 = 𝑎{ −𝐹𝑥 sin(𝜃3 ) + 𝐹𝑦 cos(𝜃3 )}𝜏2 = 𝑏{ 𝐹𝑥 sin(𝜃1 − 𝜃3 ) + 𝐹𝑦 cos(𝜃1 − 𝜃3 )} − 𝜏1

(4)

CONTROL SYSTEM FOR ACTIVE TYPE
Force control system
In the lifting motion, a force control system is applied on the assist device, as shown in Figure 6 (a). Parameterof force control system is shown in Table 1. A reference joint torque vector 𝜏 = [𝜏1 , 𝜏2 ]𝑇 is calculated from areference force vector 𝐹 = [𝐹1 , 𝐹2 ]𝑇 through Eq.4 and the torque control system is constructed as shown in
Figure 6 (b). A disturbance observer is introduced in the torque control system, where 𝑃𝜏 (𝑠), 𝑃𝜏𝑛 (𝑠) and 𝑄𝜏 (𝑠)are plant, the nominal model of the plant and low-pass filter, respectively [14]. Time constant of filter𝑇τq and nominal model 𝑇n are set to 0.005 s and 0.001 s, respectively. The output of 𝑄𝜏 (𝑠) is the estimateddisturbance, including the influence of a joint angular velocity and the influence of a parameter perturbation
between a plant 𝑃𝜏 (𝑠) and its nominal model 𝑃𝜏𝑛 (𝑠) at the same time. If these estimated disturbances arecompletely compensated, the robustness against the disturbance and parameter perturbation is highly improved.
Since a lifting motion in rehabilitation is a relatively slow motion, the influence of inertial moment and viscosity
are supposed to be ignored, so that we employ open loop force control without force feedback.

(a) Whole view of force control system.
Figure 6. Force control system.

(b) Inner torque control system.

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𝐹r
𝐹
𝜏r
𝜏
𝜃
𝐹y
𝐾
𝐾v
𝑇, 𝑇n
𝑇τq
𝐴
𝑅

Table 1. Parameter of force control system.
Reference of generation force vector
Measured force vector
Reference of generation torque vector
Measured torque vector
Joint angle vector
Force at the supported position (y axis)
Flow gain
Velocity gain
Time constant and it’s nominal value
Time constant of filter
Effective sectional area of cylinder
Radius of pulley

[N]
[N]
[Nm]
[Nm]
[rad]
[N]
[−]
[−]
[s]
[s]
[m2 ]
[m]

Force control performance
We confirm the force control system on the active type assist device. As shown in Figure 7, the assist device is
put on a force plate and the top of the device pushes up an aluminum frame fixed on the ground. In Figure 8
shows the ramp response in the proposed force control system. The blue dotted line shows the estimated force 𝐹𝑦
with Eq. 4 ( 𝐹𝑥 is set to be 0) based on the torque that is calculated with the measured pressure in the cylinder. The
black dashed line indicates the reaction force from the assist device detected by a force plate. A force generated
by the assist device roughly follows the reference. The error is mainly caused by an open loop force control. In
the lifting motion, it does not require precise force control performance, so we think the obtained force control
response shows enough in our lifting motion.

Figure 7. Experiment environment

Figure 8. Force control response

SUPPORT EXPERIMENT
Support effect in a lifting motion
We evaluated the support effect during a lifting motion using the 2 types of assist devices. In the case of active
type as shown in Figure 9 (a), the assist device is attached on subject’s right side and he performs a lifting motion
on a force plate using force control. When the touch switch on the carry case is pushed, the force control starts
supporting the subject with a ramp signal in the 𝑦 direction as the reference in Figure 6 (a). The reference velocity
in the ramp signal is set by trial and error based on the actual lift motion. In the case of passive type, theexperimental situation is the same with that in the case of active. As shown in Figure 9 (b), a subject lifts an objectof 10 kg in weight up and down without bending the elbow using the squat method. The subjects are two healthy
adult males (173±5 cm in height, 55±1 kg in weight) and four trials are performed with each subject. The raw
EMG data measured with myoelectric sensors (Oisaka Electronics Co.) placed on the dominant muscle shown in
Figure 10, are processed to IEMG with an integration time of 0.2 s.

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The 11th JFPS International Symposium on Fluid Power 2020(a) State of lifting motion.

(b) Series of lifting motions.
Figure 9. Evaluation in support experiment.

(a) Upper body.

(b) Upper legs.

Figure 10. The place of electrodes on the dominant muscles.

Figures 11 and 12 show the muscle burden of each dominant muscle shown in Figure 10, for one subject during
the lifting up and down motion with active type and passive one, respectively. The horizontal axis represents thenormalized time, which is the interval between lifting and putting off the object based on the signal from the floor
reaction force and the situation of the switch. From Figures 11 and 12, we can confirm the large decrease inmuscle burden of Rectus femoris by using both devices comparing with no device. The muscle burden of Erector
spinae is decrease, though not as much as muscle burden of Rectus femoris.
Figure 13 summarizes all the experimental results for the two subjects. The average value of IEMG with a payloadof 10 kg without the device is set as the standard (100%). From Figure 13, we can confirm that there is a large
decrease in the muscle activity ratio of Rectus femoris by using both devices with a significant difference (p<0.01).
In generally, in the squat method, a larger burden is imposing on the muscle in the upper leg compared to that in
the upper body. We can only confirm the decrease in the muscle activity ratio of the right side Erector spinae by
using the passive type device with a significant difference (p<0.01). But muscle activity of upper legs with passive
type is not decreased as much as that of active one. We think the reason is due that the passive type device becomesa resistance when the subject lifts down. In the future work, we will design a burden reduction mechanism inlifting down motion, as well as investigate the influence and effect against one side support with both devices.

(a) Erector spinae without device.

(b) Erector spinae with device.

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The 11th JFPS International Symposium on Fluid Power 2020(c) Rectus femoris without device.
(d) Rectus femoris with device.
Figure 11. IEMG with active type.

(a) Erector spinae without device.

(b) Erector spinae with device.

(c) Rectus femoris without device.
(d) Rectus femoris with device.
Figure 12. IEMG with passive type.

(a) Active type.
Figure 13. Comparison of muscle activation.

(b) Passive type.

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CONCLUSION
In this study, we developed a passive type assist device as well as active one. We conducted experiments to verify
the support effect in both of the assist device during a simple lifting motion with a payload by measuring the EMG
on the dominant muscle. In generally, in the squat method, a larger burden is imposing on the muscle in the upper
leg compared to that in the upper body. From the experimental results, we can confirm that there is a large decreasein the muscle activity ratio in Rectus femoris by using both devices with a significant difference (p<0.01) whichshows the effectiveness of the developed power assist devices. In the future, we will analyze the state of various
muscle activities in lifting motion with a musculoskeletal simulator.

REFERENCES
1. Ministry of Health, Labor and Welfare, Occupational Disease Occurrence Survey Report, 2018 (in Japanese).
2. NAMINOHIRA, K., FUJIMURA, M., Electromyographic and Kinematical Study of Human Lifting Motion -Effect ofMass of Heavy Object on Motion Methods-, Japanese society of occupational medicine, 2010, Vol.58, pp. 234-239 (inJapanese).
3. NAMINOHIRA, K., FUJIMURA, M., SHINKODA, K., Electromyographic and Kinematic Study of Lifting Motion in
Male Adults -Effects of Mass of the Object on Motion Patterns (The Second Report)-, Japanese society of occupational
medicine, 2011, Vol.59, pp. 245-250 (in Japanese).
4. FUJIMURA, M., NORA, I., The influence on lumbago in lifting weights measured by electrocardiography, Japanesesociety of occupational medicine, 2004, Vol.52, pp. 341-347 (in Japanese).
5. LI, S., NORITSUGU, T., TAKAIWA, M., SASAKI, D., Design of Wearable Power Assist Wear for Low Back Support
Using Pneumatic Actuators, Journal of Robotics and Mechatronics, 2013, Vol.7, No.2, pp. 228-236.
6. KAWAI, S., NARUSE, K., YOKOI, H., KAKAZU, Y., An Analysis of Human Motion for Control of a Wearable Power
Assist System, Journal of Robotics and Mechatronics, 2004, Vol.16, No.3, pp. 237-244.
7. NARUSE, K., KAWAI, S., YOKOI, H., KAKAZU, Y., Design of Wearable Power-Assist Device for Lower Back Support,
Journal of Robotics and Mechatronics, 2004, Vol.16, No.5, pp. 489-496.
8. MAMIYA, S., TAKAHASHI, T., UCHIYAMA, N., Design and Experimental Verification of a Pantograph-Based
Mechanism for Lower Limb Load Reduction by Compensating for Upper Body Weight, Journal of Robotics andMechatronics, 2018, Vol.30, No.1, pp. 33-42.
9. NARUOKA, Y., HIRAMITSU, N., MITSUYA, Y., A Study of Power-Assist Technology to Reduce Body Burden During
Loading and Unloading Operations by Support of Knee Joint Motion, Journal of Robotics and Mechatronics, 2016, Vol.28,
No.6, pp. 949-957.
10. NORITSUGU, T., SASAKI, D., KAMEDA, M., FUKUNAGA, A., TAKAIWA, M., Wearable Power Assist Device forStanding Up Motion Using Pneumatic Rubber Artificial Muscles, Journal of Robotics and Mechatronics, 2007, Vol.19,
No.6, pp. 619-628.
11. YOKOTA, M., TAKAIWA, M., Development of Non-Wearing Type Pneumatic Power Assist Device -Basic Concept andPerformance Evaluation-, Journal of Robotics and Mechatronics, 2020, Vol.32, No.5, pp. 1052-1060.
12. MINISTRY OF HEALTH, LABOR AND WELFARE, National Health and Nutrition Survey Report, 2017 (in Japanese).
13. CHAFFIN, D. B., ANDERSSON, G. B. J., Occupational Biomechanics, 1984, Wiley.
14. TAKAIWA, M., NORITSUGU, T., Positioning control of pneumatic parallel manipulator, International Journal ofAutomation Technology, 2008, Vol.2, No.1, pp. 49-55.

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Genaral Session | Pneumatics

Robotics and Mechatronics 1[GS1-1-01] Development of a Simple Servo-Pneumatic Three DOF Pick-And-Place
Manipulator
Chin-Yi Cheng1, Jyh Chyang Renn1, Shyang Jye Chang1, 〇Ilham Saputra1 （1.National Yunlin
University of Science and Technology）[GS1-1-02] Development of Fingertip Mechanism With Contact Point Estimation〇Kei Mikami1, Kotaro Tadano1 （1.Tokyo Institute of Technology）[GS1-1-03] Development of Outdoor Activity Assist Suit
〇Toshihiro Yoshimitsu1, Rui Matsumoto （1. Kanagawa Institute of Technology）[GS1-1-04] Examination on Attitude Control System of Hand Manipulator with
Compact Pneumatic Cylinders by E-FRIT〇Shogo Tomita1, Eiji Murayama1, Yukio Kawakami1 （1.Shibaura Institute of Technology）[GS1-1-05] Design and Fabrication of a Soft Filament-polymer Jamming Actuator〇Peng Qin1, Zhonghua Guo1, MengYu Dou1, Zhongsheng Sun1, Yan Teng1, Xiaoning Li1（1.Nanjing University of Science and Technology(China)）[GS1-1-06] Robotic Blood Vessel Mechanism for Self-Healing Function of Soft
RobotsKenjiro Tadakuma1, Shohei Inomata1, Yuta Yamazaki2, Fumiya Shiga2, Masanori Kameoka2,
MD Nahi Islam Shiblee2, 〇Issei Onda1, Tomoya Takahashi1, Yu Ozawa1, Masahiro Watanabe1,
Hidemitsu Furukawa2, Masashi Konyo1, Satoshi Tadokoro1 （1.Tohoku University, 2.Yamagata
University）©The Japan Fluid Power System SocietyGS1-1-01

The 11th JFPS International Symposium on Fluid Power 2020

Development of a Simple Servo-Pneumatic 3-DOF Pick-and-Place
Manipulator
Chin-Yi Cheng*, Jyh-Chyang Renn**, Shyang-Jye Chang***, Ilham Saputra*****Department of Mechanical Engineering
National Yunlin University Science & Technology, Taiwan(E-mail: zezheng@yuntech.edu.tw)**Department of Mechanical Engineering
National Yunlin University Science & Technology, Taiwan(E-mail: rennjc@yuntech.edu.tw)***Department of Mechanical Engineering
National Yunlin University Science & Technology, Taiwan(E-mail: changjye@yuntech.edu.tw)***Department of Mechanical Engineering
National Yunlin University Science & Technology, Taiwan(E-mail: ilhamsaputra2109@gmail.com)

Abstract. The purpose of this paper is to develop a brand new parallel-connected servo-pneumatic 3-DOF manipulator
for automatic pick-and-place purpose. The kinematic design shown in this paper is different from the commerciallyavailable DELTA or EXPT pick-and-place manipulators, which are mostly controlled by electric servomotor. The
servo-pneumatic manipulator presented in this paper consists of three single-rod pneumatic cylinders, three linear
position transducers and three pneumatic proportional directional valves. In addition to the hardware development, asoftware PID controller is also implemented. To achieve the basic movement of pick-and-place as well as two
additional path tracking controls, the closed-loop PID control scheme is utilized. Finally, various experiment results
successfully verify the validity of the overall system design.
Keywords: Pneumatics, 3-DOF Manipulator, Industry 4.0, Pick-and-Place, Automation

1.

INTRODUCTION

Industrial 4.0 was proposed by Germany in 2011. It was aimed to promote and integrate the modern industrial
technology, after-sale service and product manufacturing. Three important foundations include Intelligent Sensing
System, Internet of Things (IoT) and Big Data Analysis [1, 2]. On the other hand, pneumatic systems have thecharacteristics of fast speed, precise control and easy maintenance. Therefore, it is currently widely used in modern
automation industry. It is also believed that pneumatic control systems have been playing a very important role toimplement the Industrial 4.0. The automatic pneumatic control systems can be divided into two types, one is open
loop control scheme and the other is closed-loop control system. The former focuses on the automatic sequencecontrol. However, the fast and accurate position or velocity controls are the significant merits of the latter.
This paper is aimed to develop a brand new parallel-connected servo-pneumatic 3-DOF manipulator for
automatic pick-and-place purpose. The kinematic design shown in this paper is different from the commerciallyavailable DELTA or EXPT pick-and-place manipulators, which are generally controlled by electric servomotor.
The servo-pneumatic manipulator presented in this paper mainly consists of three single-rod pneumatic cylinders,
three linear position transducers and three pneumatic proportional directional valves. Figure 1 shows the hardware
structure of the proposed servo-pneumatic manipulator. It is worth mentioning that the design idea proposed inthis paper arises actually from the well-known 3-DOF ship or yacht simulator [3]. However, the new design istrying to rotate the structure of the simulator upside down and relocate the originally upper movable platform at
the new lower position. In the following, the design details of the kinematic hardware will firstly be outlined

2.

Hardware Design

The proposed new servo-pneumatic 3-DOF manipulator possesses a closed and parallel-connected kinematic
structure, which can further be divided into three parts. They are the main body, constraints and the movable
platform with a recording pen as shown in Fig. 1. Figure 2 shows the side-view of the manipulator without therecording pen. The most important components in the whole design are the three actuating units consisting of1

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three single-rod pneumatic cylinders with three linear position transducers. These three actuating units areinterconnected by six universal joints between the fixed upper platform and the lower movable platform at asuitable inclined angle. This inclined angle eventually determines the maximal working space of the lower
movable platform. In addition, to maintain the lower movable platform always at the horizontal attitude during
the pick-and-place movement, two necessary constrains are also interconnected between the fixed upper platform
and the lower movable platform as shown in Fig. 2. The design scheme of the constraint is depicted in Fig. 3.
Obviously, such a constraint offers only 2-DOF movements which are the slide and rotation along the axis of ballbearings as shown in Fig. 3. Furthermore, from Fig. 1, there are two constraints placed at two different verticalfaces, which will successfully limit the movement of lower platform always at the horizontal attitude.
Finally, the lower movable platform is equipped with a vertical pen located at the center position. This pen can beused to record the actual trajectory of the platform center point. However, in the future real applications, this pen
can be easily replaced by a common pneumatic suction pad or other tools to accomplish the automatic pick-andplace assembly. After real experiments, the maximal reachable working space of the proposed manipulator isfound to be W100mm x D100mm cm x H220mm. Finally, to accurately control the movement of the lower
platform, the inverse kinematic model will be used and discussed in the following section.

Figure 1. Hardware structure of the servo-pneumatic manipulator.

Figure 2. Side-view of the servo-pneumatic manipulator.

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Figure 3. Design scheme of the constraint.

3.

Inverse Kinematic Model

The inverse kinematics that maps Cartesian space into joint space is inevitably necessary for the attitude control[4, 5]. Using Cartesian coordinate, the coordinate system {B} for the upper base and coordinate system {P} forthe lower movable platform can be defined as shown in Fig. 4. Obviously, from the vector diagram shown in Fig.
4, the following Eq. (1) holds.
𝑺𝒊

𝑩

𝑷𝒊

𝑩

𝑩𝒊 ,

(1)

where the vector 𝑩𝑩𝒊 is readily known from the hardware design geometry. In addition, the vector 𝑩𝑷𝒊 can becalculated by
𝑩

𝑷𝒊

𝑻𝒗

𝑹

𝑷

𝑷𝒊 ,

(2)

where the vector 𝑷𝑷𝒊 is also known from the design geometry. The most important vector in inverse kinematic
model is the vector R, which is called the attitude orientation matrix and can be formulated by the Eq. (3) [4].
𝑐𝛾𝑠𝛾 0 𝑐𝛽 0 𝑠𝛽 1 000
1 0 0 𝑐𝛼(3)𝑠𝛾 𝑐𝛾 0𝑠𝛼 ,𝑠𝛽 0 𝑐𝛽 0 𝑠𝛼 𝑐𝛼001where cα denotes cos(α), sα denotes sin(α), and so on. The angles of pitch, roll and yaw motions at the center
point of movable platform relative to the base coordinate are denoted by γ, β and α. However, these three angles
are all equal to zero because of the two additional constraints in the kinematic design as shown in Fig. 1. Therefore,
the attitude orientation matrix can be simplified as
𝑹

𝑅 𝑅 𝑅

𝑹

1 0 0
0 1 0 .
0 0 1

(4)Finally, the length Li of every pneumatic cylinder is actually the scalar length of the vector Si. From Eq. (1), thevector Si can be rewritten as
𝑏
𝑢
𝑥̅
𝑣
𝑏
𝑦 (i = 1, 2, 3)𝑤𝑧̅𝑏Thus, the length Li of every pneumatic cylinder can be d
𝒔𝒊

(5)

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𝐿xyz(i = 1, 2, 3)(6)This is exactly the closed-form solution of the proposed parallel-connected servo-pneumatic 3-DOF manipulator.

Figure 3. Vector diagram between two working spaces.

4.

Closed-loop PID Position Control and Hardware Design

In this paper, traditional but effective PID controller is utilized to control the proposed servopneumatic 3-DOF manipulator. Figure 5 shows the control block diagram of a single pneumatic
cylinder. The schematic hardware layout of the closed-loop position control of one pneumatic
cylinder is depicted in Fig. 6. It is worth mentioning that the whole manipulator system consistsof three parallel-connected pneumatic cylinders. Thus, simultaneous control of these threecylinders is necessary. From Fig. 6, two most important components can be observed. The firstone is the proportional directional control valve (FESTO, MTYPE-5-1/8LF-010-B) and thesecond one is the potentiometer for measuring the actual position of the cylinder rod. Finally,
the PC-based controller is implemented by using the LabView software. Figure 7 shows thereal picture of the developed servo-pneumatic 3-DOF manipulator.

Figure 5. Control block diagram of a single pneumatic cylinder.

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Figure 6. Schematic hardware layout of the closed-loop position control of a single pneumatic cylinder.

Figure 7. Real picture of the developed servo-pneumatic 3-DOF manipulator.

5.

Experimental Results and Discussion

5.1 Position control of one single cylinder
In this paper, the trial-and-error approach is utilized to acquire the most suitable gains for the
PID controller. Figure 8 and Figure 9 show two experimental results of 60 mm and 120 mm
position control for one cylinder, respectively. Obviously, the position control results are
satisfactory. However, the dead-zone at the beginning of the position control is also observed.
This is chiefly because of the nonlinearity inherent in the proportional directional valve as wellas the static friction effect. The percentage steady-state position control error is less than 1%.

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Figure 8. Experimental result of 60 mm position control for one cylinder

Figure 9. Experimental result of 120 mm position control for one cylinder

5.2 Path Tracking Control
In this paper, two different paths are used to evaluate the performance of the developed servopneumatic 3-DOF manipulator. The first one is a circle and the second one is an isosceles
triangle. Figure 10 shows the experimental result of tracking a circle of diameter 48 mm. The
maximal and minimal percentage tracking error are found to be 8.3% and 0.0052%,
respectively. However, the overall average percentage tracking error is 4.2%. On the other hand,the experimental result of tracking an isosceles triangle is shown in Fig. 11. The side length of
the isosceles triangle is 100 mm. After calculation, the overall average percentage tracking
error is found to be 6.4%.

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Figure 10. Experimental result of tracking a Circle (φ: 48 mm)

Figure 11. Experimental result of tracking an isosceles triangle (side length: 100 mm)

6.

Conclusion

In this paper, a new parallel-connected servo-pneumatic 3-DOF manipulator for automatic
pick-and-place is successfully developed and realized. In addition to the basic movement ofpick-and-place, two additional path tracking controls, including the tracking of a circle and an
isosceles triangle, are also used to evaluate the performance of the developed servo-pneumatic
3-DOF manipulator. Obviously, compared to some commercially available servomotorcontrolled manipulators, the experimental percentage control errors appear in this paper arequite large. One main reason for these large errors is that the construction and assembly of thewhole manipulator system are accomplished in the laboratory. Therefore, inaccurate
manufacturing of components as well as imprecise assembly are both inevitable. In the future,
however, the developed servo-pneumatic 3-DOF manipulator may find more potential
applications in the field of automation and Industrial 4.0.
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ACKNOWLEDGMENTS
The financial supports from the Ministry of Science and Technology under grant number MOST 107- 2622- E224- 017- CC3 is greatly appreciated.

REFERENCES
1.
2.
3.
4.
5.

S.M. Song and C.D. Zhang, “Forward Kinematics of a Class of Parallel (Stewart) Platforms with Closed Form
Solutions ,,” Prof. of IEEE International Conference on Robotics and Automation, Vol. 2, 1991, pp. 2676 2681.
D. Stewart , “A Platform with Six Degrees of Freedom,” Proceedings of the Institution of Mechanical Engineers, vol.
180, part 1, no. 5, 1965, pp. 371-386.
M. H. Chiang, W. H. Chou and Y. N. Chen, “Analysis and Control of a Three-Axial Pyramidal Pneumatic Parallel
Manipulator,” Proceedings of 9th JFPS International Symposium on Fluid Power, Matsue, Japan., 2014.
C. T. Chen, J. C. Renn and Z. Y. Yan, “Experimental Identification of Inertia and Friction Parameters for Electrohydraulic Motion Simulator,” J. of Mechatronics, Vol. 21, Issue 1, 2011, pp. 1-10.
J. C. Renn, B. H. Li, “A Robot Arm Based on Parallel Pneumatic Flexible Muscle Actuators,” MATEC Web ofConferences, Vol. 123, 2017, No. 00020.

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Development of Fingertip Mechanism With Contact Point Estimation
Kei MIKAMI* and Kotaro TADANO***

Department of Mechanical Engineering, School of Engineering, Tokyo Institute of Technology
4259 Nagatsuta-cho, Midori-ku, Yokohama 226-8503, Tokyo, Japan(E-mail: mikami.k.ac@m.titech.ac.jp)**
Institute of Innovative Research, Tokyo Institute of Technology
4259 Nagatsuta-cho, Midori-ku, Yokohama 226-8503, Tokyo, Japan(E-mail: tadano.k.aa@m.titech.ac.jp)

Abstract. With the aim of achieving stable grasping with a robot hand, research has thus far been conducted on
detecting the slip and contact conditions as well as adjusting the grasping force. However, fault tolerance is anoteworthy concern because the robot hand is expected to come into frequent contact with objects when the tactile
sensor is positioned at the fingertip. Therefore, in this paper, a method is proposed for slip detection and contact
point estimation that does not require a tactile sensor. This method involves the combination of a pneumaticallydriven parallel link finger module and a fingertip mechanism using a pneumatic cylinder. Through experiments, theability to estimate the contact point position via the proposed finger mechanism is confirmed. Therefore, this method
is expected to improve failure resistance and help realize a robot hand that can grip stably.
Keywords: finger mechanism, pneumatic, grasping control, slipping, tactile sensor-less system

1. INTRODUCTION
In recent times, the concept of human–robot coexistence in environments such as factories as well as in medical
and service fields has been actively researched. Given the presence of robots in such environments, it isimportant to consider the performance and movement of the robot's hand, which comes into contact with objects
during the robot's operation; moreover, the complicated movements and grasping forces of the robot should also
be appropriately controlled. These aspects are being actively investigated [1] [2]. In the current industrial robot
system, the robot's functions are divided between grasping an object via the hand and manipulating it via thearm. If the robot is to operate while maintaining the orientation of the object once grasped, the object should bepositioned such that it can be picked up easily. This is possible in a structured environment such as a factory inthe manufacturing industry, but very difficult in an unstructured environment such as in medical and agricultural
fields. If the object can be manipulated, its orientation can be changed after it is picked up, which greatly
increases the number of applicable tasks that can be performed with the object in hand. In addition, throughcompensation for the grasping force immediately when slipping occurs, the object can be gripped and prevented
from falling down. Through measurements of the slip and compensation for the grasping force, the robot hand
can be enabled to stably grip objects even when their properties (shape, mass, coefficient of friction, etc.) are notconstant, or when an external force is applied to the object owing to a sudden stop in movement of the arm
during transportation of the robot. Therefore, during manipulation of an object in hand, it is important to be able
to detect the points at which the object is in contact with the finger.
To this end, some studies have used a contact sensor to detect the grasping state from the contact surface
information. In addition, other studies have been conducted to adjust the grasping force by detecting the slip andcontact state using a tactile sensor. Melchiorri proposed a method to control the grasping force via detection of
the translational / rotational slip using a force / torque sensor with a strain gauge and a distributed tactile sensor[3]. However, despite their usefulness, these methods require the coefficient of static friction to be known.
Maeno et al. proposed an elastic finger for detection of distribution, using a strain gauge with a curved surface,
and they showed that any object with an unknown weight and friction coefficient can be lifted at any speed[4][5][6]. Their studies have employed tactile sensors to detect the slip. However, fault tolerance is a key
concern because the robot's hand is expected to make frequent contact with objects when the tactile sensor isfixed at the fingertip. In addition, when a large number of sensors and distributed tactile sensors are mounted on
the robot hand, issues concerning wiring processing and amplifier placement are presented. The higher the
resolution is, the longer it takes to acquire sensor information and process data, which becomes a constraint
during real-time control.
In this context, it is also possible to employ vision to obtain non-contact slip and contact information [7][8].
However, the processing speed depends on the frame rate of the vision camera. In addition, high resolution isrequired to detect minute displacements. Even when a high-speed camera is used, the placement of the camera
becomes a problem, as space is often restricted around the hand, and a fixed camera may give rise to a camera

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blind spot. It is also often necessary for adjustments to be made each time depending on the surrounding
environment in terms of factors such as brightness.
To address these drawbacks, a pneumatically driven finger module is proposed in this paper. By means ofdriving the five-bar link using two pneumatic cylinders, the grasping force can be estimated without employing
a force sensor, and a high grasping force-weight ratio can be achieved. Moreover, tactile sensor-less slip
detection and contact point estimation are possible through the combination of the finger module and thefingertip mechanism via a pneumatic cylinder. By means of monitoring the contact point oscillation within thecontrol cycle, it is expected that the slip can be detected at high speed.
When an external force is applied to the pneumatic cylinder with the compressed air supply valve closed, theinternal pressure of the cylinder changes. The contact point can then be obtained by means of the estimatedgrasping force from the finger module and the variation in internal pressure of the cylinder. As the proposedmechanism does not involve a tactile sensor, it has an advantage in terms of the fault tolerance not being an
issue.
The remainder of this paper is organized as follows. In Section 2, an outline of the proposed fingertip
mechanism and contact point estimation method are presented. In Section 3, the control system of the fingertip
mechanism and finger module is explained. In Section 4, the experimental results are presented. Finally, Section
5 summarizes the study and presents the concluding remarks.

2. FINGERTIP MECHANISM FOR CONTACT POINT ESTIMATION
2.1 Overview
Figure 1 shows the configuration of the developed fingertip mechanism and each variable thereof, and Figure 2
shows the layout of the proposed fingertip mechanism with the finger module developed in this study. The
fingertip mechanism employs a pneumatic cylinder. The supply valve is opened, and compressed air is sent to
one of the air chambers through the tube. Then, the supply valve is closed. The change in internal pressure
within the pneumatic cylinder is measured using a pressure sensor. The other air chamber is open and underatmospheric pressure. The grasping force Fgrip is the vertical force on the grasping surface of the tip link. The
state corresponding to Fgrip = 0 N is defined as the initial state. When Fgrip and the position of the contact point
are changed, the load applied to the pneumatic cylinder changes, followed by the cylinder length ls.

Fs

mR

ms

Fgrip
Tip link

ls

θs

ϕs

lw

msb

otip

mL

ObjectFigure 1. Finger module configuration
Pneumatic cylinder

Pneumatic cylinder (Finger module)
Potentiometer

Proposedfingertip mechanism

50.6
Five-bar link
Tip link
100

46
40

Figure 2. Layout of proposed fingertip mechanism with finger module2

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2.2 Contact Point Estimation Method
This section describes the proposed contact point estimation method via the fingertip mechanism. The torquearound the fingertip joint otip is represented by the following relationship.
F g rip l w = F s sin θ s ⋅ m s .

(1)Here, lw is the distance between the contact point at the object and otip, Fs is the reaction force of the pneumatic
cylinder, θs is the angle between the pneumatic cylinder and fingertip link, and ms is the distance between therod end of the pneumatic cylinder and otip as shown in Figure 1. Fs is obtained as follows:
Fs = P A .

(2)Here, P is the internal pressure (gauge pressure) of the pneumatic cylinder, and A is the area subjected to thepressure. θs can be obtained from the cosine theorem using the following equation.
2
 m s2 + l s2 − m sb

.
2
m
l
s s



θ s = co s − 1 (3)Here, msb is the distance between the base of the pneumatic cylinder and otip. Assuming an isothermal change,
the estimated cylinder length lˆs can be obtained from the changes in the internal pressure and volume of thepneumatic cylinder according to Boyle's law as follows:
V −V lˆs = l s 0 −  0

A 

V =

P0 + 1 0 1 .3 2 5
V0 .
P + 1 0 1 .3 2 5

(4)

(5)

Here, V is the volume of the cylinder and pipeline together; and P0, V0, and ls0 are respectively the pneumatic
cylinder internal pressure, volume, and cylinder length under no load. On substitution of Eq. (2)–(5) into Eq. (1),the estimated value of lw is expressed as follows.
2

 2 

 P + 101.325  V0 2 ms +  ls 0 − 1 − 0  − msb  
P + 101.325  A 
PA


m .

lˆw =
sin  cos −1 
 s
Fgrip
2msls










(6)

Fgrip is obtained from the external force Fext at the tip of the finger module and the fingertip link angle φs. The
fingertip link angle φs can be obtained from the cosine theorem using the following equation.
2
 m s2 + m sb
− l s2  .
2 m s m sb


φ s = co s − 1 (7)As this calculation process does not require consideration of parameters related to the properties of the object, it
is useful even in an unstructured environment.
The external force Fext can be measured directly using a force sensor; however, in the case of the finger module
proposed herein, Fext can be estimated from the differential pressure of the pneumatic cylinder and thekinematics model without a force sensor being required. This approach can potentially reduce costs and improve
fault tolerance. In this study, the estimated value Fˆ ext , which is described later, is considered as the external
force at the tip of the finger module.

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2.3 Proposed Fingertip Mechanism
Figure 3 shows the developed fingertip mechanism with finger module. The total mass is 0.29 kg. Thismechanism employs a pneumatic cylinder (SMC, CP2B4-5D). The drive cylinders of the finger module are thelow-friction pneumatic cylinders (SMC, CJ2XB16-30Z) and were controlled via a five-port pneumatic
proportional valve (FESTO, MPYE-M5-B) with an operating band of approximately 100 Hz. A linear
potentiometer was used to detect the position of the cylinder rod, and the intake and exhaust of the pneumatic
cylinder were controlled via a pneumatic solenoid valve (SMC, VQZ312-5M1-C6). The supply pressure to the
proportional valves was set to 500 kPa, and the initial supply pressure of the pneumatic cylinder was set to 100
kPa. A semiconductor pressure sensor (SMC, PSE540) was used to measure the pressure of each port and thepneumatic cylinder.

3. STRUCTURE OF CONTROL SYSTEM
Figure 4 shows a block diagram of the grasping force compensation control and the compliance control systemof the finger module. It can be applied to a pneumatically driven surgery support robot that estimates andcontrols external force without a force sensor, and also to the proposed mechanism. The reference force Fref was
generated from the position error of the tip and transformed into the reference driving force Fcyref for eachcylinder using the Jacobian J a . Moreover, Fcyref was controlled using a PI controller with a feedback force Fcy ,calculated from the pressure difference in the cylinder. The control signal u was sent to the pneumatic
proportional valve, and the force was generated by charging the compressed air into the actuator.
For a grasping task to be executed successfully, compliance and adaptability of the robot hand must be ensured
for stable grasping. In this study, the method proposed by Tadano and Kawashima [9] was adopted for thecontrol system. This control system has been applied to pneumatically driven surgery robots that perform
external force estimation and force control without force sensors, and it is suitable for the proposed finger
module.
Pneumatic cylinder

Pneumatic cylinder (Finger module)

PotentiometerFive-bar link
Tip Link

Fingertip jopint

Figure 3. Overview of proposed fingertip mechanism with finger module
X

F.K.

Bs

Xcmd

+

−

K

− Fref+Fcyref− R

Ja

T

−

+

Fˆ extrefcy − LF

P

Contact point
Estimator (Eq.(6))+−

K ap + K ai s

K ap + K ai s

u1

u2

Valve R

Cylinder R

Fcy − R
FingerMechanism

Valve L

Cylinder Lq

Fcy − Llˆs

Figure 4. Block diagram of control system

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The fingertip was designed as per the motion characteristics described in Eq. (8), where B is the viscosity and K
is the stiffness.
BFˆext =  x
 00   xɺ   K x+
B z   zɺ   00   x cm d − x res  .
K z   z cm d − z res (8)

We constructed a control system to drive the developed finger module. An ADC board was used to capture thesignals of the potentiometer and pressure sensor. Additionally, a voltage signal was input from the DAC board
to the proportional valves. The control cycle was set to 1 ms.

4. EXPERIMENTAL RESULTS AND DISCUSSION
4.1 Experiment for Assessing Maintenance of Cylinder Internal Pressure
Accurate estimates via the proposed method cannot be obtained if an air leak is present. Therefore, it wasnecessary to confirm the airtightness of the experimental system. After the solenoid valve was closed under theinternal pressure of the pneumatic cylinder, the cylinder rod was moved 10 times with a full stroke within 60
seconds. Then, ls was measured using a potentiometer attached as shown in Figure 5.
Figure 6 (a) shows the measured value of the cylinder internal pressure, and Figure 6 (b) shows the measured
value of ls. Before and after the experiment, the internal pressure was held at 100 kPa. The amount of air leakage
from the experimental system was deemed small enough to be neglected. However, in practical use cases, theinternal pressure of the cylinder may gradually decrease when the cylinder is operated continuously for a longtime. However, there is a period of time when the robot is not in contact with the object during the work cycle.
At this time, if the internal pressure can be held for approximately 60 seconds, the solenoid valve can be opened
so as to restore the cylinder pressure when the fingertip is not in contact with the object.

4.2 Experiment for Contact Point Estimation
As part of the study, we conducted an experiment to estimate the contact point using the developed fingertip
mechanism and finger module. The object—an arc-shaped resin—was positioned on a linear stage so as to be
movable horizontally. The displacement of the linear stage was measured using a potentiometer. An overview ofthe experimental system is depicted in Figure 7. The experimental parameters are listed in Table 1.
After the fingertip link was pressed against the object, the linear stage was moved to the left, as shown in Figure
7. Figure 8 (a) shows the internal pressure of the cylinder, and Figure 8 (b) shows the comparison between themeasured cylinder length ls (dotted line) and the estimated cylinder length lˆs (solid line). The displacement of
the initial length of the pneumatic cylinder from 14.5 mm to 11 mm could be estimated from the pressurechange.

Pneumatic cylinderls

Tip link

Potentiometer

Figure 5. Potentiometer for air cylinder of fingertip mechanism

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120

P [kPa]

110
100
90

Pressure

80
70
60
60
0

10

20

30
40
Time[s](a) Cylinder internal pressure

50

60

16
14

ls [mm]12108

ls (Measured)

6
4
2
60
0

00

10

20

30
40
50
60
Time[s](b) Cylinder length
Figure 6. Results of experiment for assessing maintenance of cylinder internal pressure

Tip position of the finger module
Contact point

ObjectPotentiometer

Liner stage

Figure 7. Overview of experimental system

ls0
V0
P0

Table 1. Parameters of experimental system
14.5 [mm]K
10.0 [N/mm]
967.6 [mm3]B
0.01 [Ns/mm]
100 [kPa]

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Presuurer[kPa]

120
110
100
90
80
Pressure
P5[kPa]

70
60
0

2

4

6

8

10

Time[s]Cylinder length[mm](a) Result of measuring cylinder internal pressure
20

15
10

ls (Measured)系列ls
ˆl (Estimated)
系列
ls_m…
s

5
0
0

2

4

6

8

10

Time[s](b) Results of measuring cylinder length
Figure 8. Experimental result for cylinder length estimation

Figure 9 (a) shows the movement distance of the object measured using the potentiometer (Figure 7). Figure 9(b) shows the comparison between the measured contact point lw and the estimated contact point lˆw . The
measured contact point lw was obtained via geometric calculations based on the moving distance of the linear
stage, cylinder length, and fingertip joint position, which were obtained from each potentiometer. The measured
contact point lˆw was obtained using Eq. (6)The error between the measured and estimated values was approximately 3 mm or less, which was reasonable.
As the estimated cylinder length adequately agreed with the measured value, the estimated grasping force usedin the estimation calculation was supposed as the cause of the error. Although there were improvements in thegrasping force estimation, changes in the contact position could be detected. Therefore, the estimated the contact
position can be used for slip detection.
In the case of detection of slip, it was assumed that the slip could be suppressed by compensating for Fgrip
proportionally to the amount of fluctuation of lw from the no-load state. By means of calculating thecompensation force based on the amount of fluctuation lw, the grasping force can be appropriately compensatedeven if there is an error in the absolute value of lw. The static friction and hysteresis of the mechanism were
determined to be the main factors affecting the error of the grasping force estimation. Ultimately, by means ofimproving the accuracy of the friction model, the error in lw can be minimized.

CONCLUSION
In this study, we proposed a finger mechanism employing a pneumatic cylinder for contact point estimation. It
was experimentally confirmed that the proposed finger mechanism could yield an accurate estimate of theposition of the contact point without the requirement for a tactile sensor. Furthermore, by means of combining
this mechanism with the finger module developed in this study, a low-cost robot hand with excellent fault
resistance can be realized. Overall, the proposed finger mechanism is an elemental technology prototype of therobot hand, whereby stable grasping can be achieved.
In future work, we will consider the Fgrip compensation method for stable grasping. In addition, we plan to
manufacture a multi-finger hand based on the proposed finger mechanism and evaluate its functionality.

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Moving distance[mm]

15

10

5
MovingDistane
distance

0
0

2

4

6

8

10

Time[s](a) Moving distance of the object

Contact point [mm]

40
30
20

lw (Measured)
系列
lwLPF
ˆl (Estimated)
接点…
系列
w

10
0
0

2

4

6

8

10

Time[s](b) Results of estimating contact point
Figure 9. Experimental result for contact point estimation

REFERENCES
1. KAWASAKI, T., et al., Dexterous Anthropomorphic Robot Hand with Distributed Tactile Sensor: Gifu Hand II,
IEEE/ASME Trans. on Mechatronics, Vol.7, No. 3, pp.296-303 (2002)
2. MIZOGUCHI, Y., et al., Development of Intelligent Robot Hand Using Proximity, Contact and Slip Sensing, Trans. Soc.
Instrum. Control Eng., Vol.46, No.10, pp.632-640 (2010) (in Japanese)
3. C. Melchiorri, Slip Detection and Control Using Tactile and Force Sensors, IEEE/ASME Transaction on Mechatronics,
vol.5, no.3, pp.235-243, 2000.
4. MAENO, T., et al., Control of Grasping Force by Detecting Stick/Slip Distribution at the Curved Surface of Elastic
Finger, Proceedings of IEEE International Conference on Robotics and Automation, pp. 3896-3901, 2000
5. YAMADA, D., et al., Artificial Finger Skin having Ridges and Distributed Tactile Sensors used for Grasp Force Control,
Journal of Robotics and Mechatronics, Vol. 14, No. 2, 2002, pp. 140-146
6. MAENO, T., et al., Grip Force Control by Detecting the Internal Strain Distribution Inside the Elastic Finger Having
Curved Surface, Trans. JSME C, Vol.64, No.620, pp.1258-1265 (1998) (in Japanese)
7. IKEDA, Y., et al., Grip Force Control for an Elastic Finger Using Vision-Based Incipient Slip Feedback, Proceedings ofIEEE/RSJ International Conference on Intelligent Robots and Systems, pp.810-815, 2004
8. IKEDA, A., et al., Grip Force Control of the Elastic Body based on Contact Surface Eccentricity During the Incipient Slip,
Journal of Robotics Society of Japan vol. 23, no. 3, 2005, pp. 337-343. (in Japanese)
9. TADANO, K., et al., Development of a Master Slave Manipulator with Force Display using Pneumatic Servo System for
Laparoscopic Surgery, Proceedings of 2007 IEEE International Conference on Robotics and Automation, pp. 947–952,
2007

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Development of Outdoor Activity Assist Suit
Toshihiro YOSHIMITSU *, Rui MATSUMOTO ** Department of Robot mechatronics, Faculty of Engineering
Kanagawa Institute of technology
1030 Shimo-oghino,Atsugi ,Kanagawa 243-0292(E-mail: yosimitu@rm.kanagawa-it.ac.jp)

Abstract. In recent years, loss of lives and rescue has increased for nature disaster. In rescue operations what active“long hour”, “outdoor” and “hill”, is broken out exhaustion and injury. Then, we invented to Assist Suit for outdoor
activities because it’ll be effective in preventing injury and tired. Assist Suit we invented uses a pneumatic cylinder
different of a lot of Assist Suit including HAL the representative Assist Suit made by Cyber Dyne use an electricmotor helped joint torque. Using pneumatic cylinder for Assist Suit, it has advantaged at weight and long houractivities we thought. In this study, we putted stress to making control system for Assist Suit and inserting a springelement in power source experimentally. The purpose of this study is to develop a lightweight assist suit actuatorsystem with lower joint motion range and suitable for outdoor activities.
Keywords: Pneumatic, Power assist, Ejector, Regenerator

INTRODUCTION
In recent years, natural disasters such as eruptions have caused an increase in human damage and rescue
activities. In rescue activities, there are many cases in which "long hours", "outdoors", "rough terrain", and"sloping areas" are performed, so that the load applied to the lower limbs increases. It is feared that the rescuer
may suffer fatigue and injuries due to burden on lower limbs.
The mechanisms used in indoor assist suits are classified as either "there is a small degree of freedom and there
is a function to support the direction in which assistance is needed" or "the joint has some flexibility but nofunction to support the joint" It is. On terrain and sloping terrain, it is necessary to implement conflicting effects
such as free movement that does not hinder the activity and assisting the movement of the joint to prevent
injuries and sprains. In this study, we develop an actuator system for assist suits that has joint range of motion
and reduces the weight of the mechanism and is suitable for outdoor activities.

Outdoor Activity Assist Suit
In situations such as mountain rescue or landslide disasters where work must be continuously performed on
uneven or sloping terrain, we develop a power assist suit for lower limbs for the purpose of reducing physicalload, extending activity time and preventing injuries to the rescuer himself. The functions required for the lower
limb assist suit include weight reduction, lower limb burden on slopes, prevention of injuries, and long-term
activities.
Figure 1 shows the outline of the outdoor activity assist suit created in this study. The impact applied to theknee joint is greater when descending than when climbing, which may cause fatigue and injuries due to it.
The aim of outdoor activity assist suit is to absorb the impact applied to the knee joints when descending a slopewith an assist unit such as a pneumatic cylinder. Specifically, the rod speed of the cylinder controls the flow rateof compressed air discharged from the head side. It is characterized by aiming for shock absorption.

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Figure 1. Outdoor Activity Assist Suit

Table1 Assist Suit Elements
Size [mm]WeightPower Source

973×475×266
7.1kg
Pneumatic Cylinder(φ32)

Knee range of motion

120Degrees

REGENERATIVE ACTUATORSThis assist suit is a passive actuator that absorbs the impact force applied to the knee joint when going
downhill. In previous research, the impact force was absorbed by controlling the opening of the speed controller
with a servomotor for flow control. The speed controller controls the outflow speed by the throttle mechanism,
but the energy is dissipated as resistance and cannot be used.
In this study, we use an ejector instead of a speed controller for flow control to absorb the impact force whendescending and to regenerate energy when descending.
EjectorActuator

P

E
V
Check
Valve

Pressure
Gauge

TANK

Fig.2 Energy regeneration system

ACTUATOR REGENERATIVE FORCEThe energy regeneration by the ejector includes an ejector cycle circuit. In this study, the vacuum pressure
generated by the ejector is used as regenerative energy to drive the actuator while controlling the speed duringhigh-speed movement of the cylinder by the throttle of the ejector. In order to confirm the effectiveness of thismechanism, two types of ejectors of different sizes, cylinder speed-load experiment, cylinder operation-vacuum
pressure experiment, were performed and the effectiveness was confirmed.
The experimental results are shown in Fig. 3,4. The flow rate was controlled by the nozzle of the ejector, andthe load increased as the effective area of the ejector was smaller, and the cylinder speed was faster.
As an energy regeneration experiment, a small pressure vessel (capacity: 18.0 cc) was connected to theejector, and the number of reciprocations of the cylinder and the pressure in the pressure vessel were measured.
The experimental results show that the smaller the effective area of the ejector, the higher the energy
regeneration effect.

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180

ZH05B

160

ZH07D

140

ZH10D

Load[N]

120
100
80
60
40
20
0
Fast

MidlleCylinder pushing speed

Slow

Fig.3 Actuator load experiment
0
-5

1st

5th

10th

15th

Pressure[kPa]

-10
-15
-20
-25
-30
-35
-40

Cylinder push-in count
ZH10D

ZH07D

ZH05B

Fig.4 Ejector vacuum pressure experiment

CONCLUSION
In this study, we proposed a mechanism using a pneumatic cylinder and an ejector as an actuator for an
outdoor activity assist suit that can save energy and be active for a long time. The mechanism proposed by thisstudy uses a throttle of a vacuum generation mechanism to absorb the impact of a physical load during outdoor
activities on a downhill, and to reduce the vacuum pressure generated when compressed air passes through theejector. The effectiveness of the system that can be used secondarily by accumulating it in the tank was
confirmed.

REFERENCES
Yoshimitsu T.，Shoji T., DEVELOPMENT OF POWER ASSISTING SUIT FOR ASSISTING RESCUE CREW , Proc. of
jfps2017 , 2D32 (2017)
2. Sankai Y.，Kawamura Y., Study on Hybrid Power Assist System HAL-1 for Walking Aid using EMG， JSME
Proceedings of the Ibaraki Lecture Meeting, pp.269-270(2000)
3. Sano K.， Yagi E.：Study on a Power Assist System for Overhead Harvesting / Slope Walking, Proc. of 54th AutomaticControl Alliance Lecture , pp.827-829 (2011)
4. TAKEUCHI .H, et al., High Efficiency "EJECTOR CYCLE" Refrigerator ，Refrigeration 79(920), pp. 431-432(2004)

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Examination on Attitude Control System of Hand Manipulator with
Compact Pneumatic Cylinders by E-FRIT
Shogo TOMITA*, Eiji MURAYAMA** , Yukio KAWAKAMI**
Graduate School of Science and Engineering, Shibaura Institute of Technology
307 Hukasaku, Minuma-ku, Saitama, 337-8570 Japan(E-mail : md20061@shibaura-it.ac.jp)**
Faculty of Science and engineering, Shibaura Institute of Technology
307 Hukasaku, Minuma-ku, Saitama, 337-8570 Japan*

Abstract. Pneumatic system has some merits that being strong against overload and safe for humans because it drivesby air, a compressible fluid. Taking advantage of that merits, some research about robots that are used pneumatic
actuator have been conducted. In the context of that, we have developed hand manipulator which is driven by compact
pneumatic cylinder. Although pneumatic actuators are more difficult to control precisely than electric motor usedconventionally, for this problem, we are studying the application of control using E-FRIT (Extended Fictitious
Reference Iterative Tuning), which is a PID gain adjustment method based on experimental data. In this paper, weshow the result of applying controller design with E-FRIT to developed manipulator.
Keywords: Hand manipulator, Pneumatic cylinder, E-FRIT, PID control, Control system

INTORODUCTION
Today, robots are intervening many of fields such as not only industry but also, care, entertainment, etc. Whilethe distance between human and robots will be close, it will be more required that we develop the robot which issafe to humans. That kind of field of research is named “soft robotics” and pneumatic actuator is collecting a lotof attention as a safe actuator [1] [2].
There are some studies of manipulator with pneumatic actuator. In generally, using pneumatic rubber artificial
muscle [3], using bursiform actuator which drive putting in and out air [4], and so on. Our team focuses on using
compact pneumatic cylinder at robot hand’s joint. Because pneumatic cylinder drives linearly, we have to convert
the linearly motion to rotational motion to adjust robot hands, on the other hand, there is a merit that it can becontrolled more easily than other pneumatic actuators because of its simple construction.
We have continued research of manipulator using pneumatic cylinder. In previous researches, some control
methods were suggested to pneumatic robotics system (e.g. Simple Adaptive Control, Sliding Mode Control) [5].
They achieved results, but there is a problem that it takes time and effort to identify the system. Also, themanipulator didn’t have reproducibility because of hardware problems. So, we made a new manipulator which
can improve the hardware problems, and we are trying to apply E-FRIT, one of the PID gain adjustment methods
to force control and attitude control. We achieved force control accuracy of ±0.3[N] by using estimated force inprevious research (Figure 1)[6], so in this study, we focus on the attitude control.

Figure 1. Achievements about force control of manipulator[6]1Copyright (C) JFPS All Rights Reserved.

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MANIPULATOR OVERVIEW
Figure 2 shows overview of developed manipulator. It has three joints as with human and it is driven by compact
pneumatic cylinder. Table 1 shows the pneumatic cylinder’s specifications. In this manipulator, slider levermechanism makes it possible to convert linearly motion to rotational motion. Figure 3 shows that mechanism
overview and Figure 4 shows that manipulator’s range of motion. At the mechanical part, we use bearings andbush made of resin to smoothly convert.
This manipulator has potentiometers and pressure sensors to control attitude and force. It is driven by pneumatic
cylinder, so we can control the force without using force sensors. If we try to control the force using force sensors(etc. strain gauge), it may not be able to do because of the shape of the object that the sensors do not touch (Figure5), so force control using pressure difference in the pneumatic cylinder do not depend on the shape of object.
Compact Pneumatic Cylinder

22[mm]

19[mm]
36[mm]

56[mm](a) Pneumatic cylinder size

55[mm]

40[mm](b)Used manipulator in this study

Figure 2. Developed manipulator

Table 1. Specification of compact pneumatic cylinder
Manufacturer
KOGANEI(Custom order)
ActionDouble acting
Fluid
Air
Bore size
10[mm]
Maximum operating pressure
1.0[MPa]
Minimum operating pressure
0.03[MPa]
Standard Strokes
20[mm]
Weight23.9[g]Linear Motion

Rotational Motion
Figure 3. The mechanism of manipulator’s joint

90[deg]110[deg]
90[deg]Figure 4. Manipulator’s joint angle

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Not touch to object

Figure 5. Manipulator’s state when grasping (not touch force sensor)

THE EXPERIMENTAL DEVICE OF MANIPULATOR
Figure 6 shows the pneumatic circuit of manipulator used in this study. The manipulator has three kinds of sensor,
potentiometers, pressure sensors, and load cells. Potentiometers are used for feedback of manipulator’s joint angle
to control its attitude. Pressure sensors are used for feedback of pressure inside the pneumatic cylinders to control
manipulator’s joint torque.
In this manipulator, the control targets are electro-pneumatic regulators which are connected compact pneumatic
cylinders at manipulator’s joints. Electro-pneumatic regulator is proportional control valve, and it continuously
control air pressure in proportion to the input electric signals. Figure 7 shows the response of electro-pneumatic
regulator (SMC: ITV0051-3MS). We control them in order to control the manipulator’s attitude and joint torque(it’s same to control manipulator’s gripping force). Incidentally, this time, we control air pressure only head-sideof cylinders. Rod-side pressures are fixed about 0.09[MPaG] (that value is electro-pneumatic regulator’s output
pressure when input 1[V]).
We used MATLAB/Simulink🄬🄬 to design controller and implemented using Simulink Real-TimeTM.

Figure 6. Pneumatic circuit used in this study

0.08991(a) Relationship between Voltage and Pressure(b) Step Response
Figure 7. Electro-pneumatic regulator response3

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PID GAIN TUNING BY E-FRIT [7]The manipulator controls attitude and force. For each, we adopted PID control about control method. PID control
refers only target value and output value so that we can control the system if we don’t know the internal model ofthe plant, but it has demerit that the difficulty of gain tuning. So in recent years, method of auto gain tuning isunder intense investigation. E-FRIT is one of the suggested method of auto gain tuning. It doesn’t have to constructsimulation models. It needs only stable closed loop data and settling time we specify.
Eq.(1)-Eq.(3) are some parts of relational expressions used in E-FRIT. Eq.(1) shows expression PID controller,
Eq.(2) shows expression of reference model, and Eq.(3) shows variable which it optimize. In Eq.(2), the value ofn is 2 if the controller is PID control or PI-D control (if you use I-PD control, the value of n is 3). T99 is settling
time which user can determine arbitrarily. E-FRIT optimize L at Eq.(2) and PID parameter at the same time.
𝐶𝐶(𝑠𝑠) = 𝐾𝐾𝑝𝑝 +𝑀𝑀(𝑠𝑠) =𝐾𝐾𝑖𝑖+ 𝐾𝐾𝑑𝑑 𝑠𝑠
𝑠𝑠

(1)

1
𝑇𝑇99
𝑒𝑒 −𝐿𝐿𝐿𝐿 , 𝜏𝜏 =𝑛𝑛(𝜏𝜏𝜏𝜏 + 1)
4.4𝑛𝑛0.6(2)(3)∅ = �𝐾𝐾𝑝𝑝 , 𝐾𝐾𝑖𝑖 , 𝐾𝐾𝑑𝑑 , 𝐿𝐿�

ATTITUDE CONTROL
We performed attitude control experiment on the manipulator’s second joint using E-FRIT explained above.
First, we performed a proportional control closed loop experiment (only P control) as preliminary experiment for
E-FRIT. Figure 8 shows the result. It is the step input response which reference value is 45[deg].
Table 2 shows the result of E-FRIT calculated gain. Based on this result, we tuned PID gain. Figure 9 shows thestep input response as the target value is 45[deg] using the tuned gain. According to this result, we were able toconfirm that manipulator’s joint angle followed reference value smoothly and overshoot was not seen. It is thoughtthat E-FRIT is valid method for manipulator which drives compact pneumatic cylinder.
Figure 10 shows the result of step response which target is 60[deg] using the same gain as the target value is45[deg]. It can be seen relaxation oscillations around target value. We found from these results that we have totuning the gain every time that we change the target value because it has low robustness. There is a small difference
about delay time in Figure 9 and Figure 10, this is due to the backlash of the manipulator.

Kp = 2.60e-2(a)Process variable (manipulator’s angle)(b) Manipulated variable(electro-pneumatic regulator)
Figure 8. Preliminary experiment result(P control)
Table 2. E-FRIT calculated result

Kp
Ki
Kd

2.792e-22.693e-2
1.000e-3

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Figure 9. Step response as the target value is 45[deg] (PID control)

Relaxation Oscillation

Figure 10. Step response as the target value is 60[deg] (same gain in Figure 9.)

SUGGESTION DISTURBANCE OBSERVER FOR ATTITUDE CONTROL
In attitude control, there is a problem that low robustness to the target value. It is thought that the cause of some
disturbances such as frictions. To achieve higher control performance, we thought to apply disturbance observer
for attitude control system. Generally, positioning control performance of pneumatic servo system is related to
the nonlinear friction characteristics, and there are some previous researches that disturbance observer used ascompensation [8] [9].
To design disturbance observer, we have to identity the system of control target to obtain nominal model. To
identify the system is a very difficult task, so we lost the advantage of E-FRIT that we can significantly reduce
the time and cost required for controller design. There, we calculate the nominal model to use model matching of
E-FRIT.
In feedback control, the transfer function P(s) from input to output can be expressed as Eq.(4).
P(s) =𝐶𝐶(𝑠𝑠)𝐺𝐺(𝑠𝑠)1 + 𝐶𝐶(𝑠𝑠)𝐺𝐺(𝑠𝑠)(4)

In this control system, C(s) is PID controller calculated by E-FRIT, and G(s) is manipulator’s plant model which
we want to know. E-FRIT defines closed loop reference model as M(s) showed Eq.(2). If E-FRIT satisfy themodel matching, we can think that Eq.(2) is equal to Eq.(4) and Eq.(5) is established. We can estimate the plantmodel to solve Eq.(5) for G(s).
𝐶𝐶(𝑠𝑠)𝐺𝐺(𝑠𝑠)1𝑒𝑒 −𝐿𝐿𝐿𝐿 =(𝜏𝜏𝜏𝜏 + 1)2
1 + 𝐶𝐶(𝑠𝑠)𝐺𝐺(𝑠𝑠)Copyright (C) JFPS All Rights Reserved.

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Figure 11 shows the result of E-FRIT model matching. In this result, there is less difference between actual
manipulator’s response (blue line) and E-FRIT designed model’s response (red line). Therefore, we consideredthat it satisfied model matching, and we designed disturbance observer from the result.
Figure 12 shows that the control system of manipulator’s attitude controller with disturbance observer. The inverse
model of nominal model is non proper, so we add Butterworth filter (F) to change to proper. Also, we add the low
pass filter (L) in estimated disturbance because potentiometer contains measurement noise.
By the use of disturbance observer made from E-FRIT, we performed manipulator’s attitude control experimentwith inputting disturbance. Manipulator’s joint angle is made to follow a target value (45[deg]), and 10 seconds
later, we input 0.5[V] to electro-pneumatic regulator’s manipulated voltage as disturbance.
Figure 13 shows the result of comparison with disturbance and estimated disturbance. From this result, we foundthat the disturbance observer made by E-FRIT can roughly estimate disturbance. Figure 14 shows actual
manipulator response. It can be seen that the response with using disturbance observer is less affected by
disturbance and it more quickly converges to the target value than not using.

Figure 11. Model matching result by E-FRIT

Figure 12. Disturbance observer

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Figure 13. Comparison input disturbance and observed disturbance(a) Whole response(b)Response around disturbanceFigure 14. Comparison of manipulator’s response to disturbances

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CONCLUSION
In this study, we developed hand manipulator which is driven by some compact pneumatic cylinders and assess
about attitude and force control performance. In previous research, we could achieve a certain result about force
control, we focused on attitude control in this study. We designed the attitude control system by PID control whose
gain tuned by E-FRIT that one of the method of experimental data auto gain tuning. We could design it to reachthe target value quickly and no over shoot, but we found the problem that the control performance deteriorates
when we change the target value. There is a need to tuned suitable PID gain every time we change the target value,
so we have to another approach to design the control system.
In order to improve the control system of attitude, we considered apply disturbance observer. In usual, we haveto identify the system because we need to know the plant nominal model to design disturbance observer, but wefocused on the model matching of E-FRIT and we devised that to derive the plant nominal model from result of
E-FRIT calculation process.
We made a manipulator’s model as described above and built the disturbance obsever. It could estimate
disturbance to some extent, and be able to improve the attitude control performance of the manipulator.
In future, we will evaluate the performance of the manipulator as a whole, not just only one joint.

REFERENCES
1. Uchiyama, M., Soft Robotics, Journal of the Robotics Society of Japan, Vol.17, No.6, 1999, pp.756-757.
2. Takaiwa, M., Status and Prospects of Pneumatic Systems, Journal of the Society of Instrument and Control Engineers,
Vol.56, No.4, 2017, pp.241-242.
3. Yasunori, O., Development of a Human Safe, Multi-Degree-of-Freedom Robot Arm Technology using Pneumatic Muscles,
Robotics Society of Japan, Vol.28, No.3, 2010, pp.302-310.
4. Nishioka, Y., Pneumatic Soft Robot Arm, Journal of the Robotics Society of Japan, Vol.33, No.9, 2015, pp.10-13.
5. Murayama, E., Model-Based Design Approach for Bipedal Robot, Hydraulics and pneumatics, Vol.55, No.3, 2016,
pp.23-29.
6. Tomita, S., Development of Hand Manipulator with Compact Pneumatic Cylinder, The Proceeding on Autumn Conferenceof Japan Fluid Power System Society, 2020, p.21.
7. Kano, N., Actual use of operational data in the process industry, IEEJ Transactions, Vol.131, No.4, 2011, pp.710-717.
8. Noritsugu, T., Design of Pneumatic Positioning Control System Using Disturbance Observer, Transactions of the Societyof Instrument and Control Engineers, Vol.31, No.1, 1995, pp.82-88.
9. Iwasaki, M., Optimal Control of Pneumatic Cylinder with Disturbance Observer, Journal of the Japan Hydraulics andPneumatics Society, Vol.27, No.6, 1996, pp.769-802.

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Design and Fabrication of a Soft Filament-polymer Jamming Actuator
Peng QIN, Zhonghua GUO, Mengyu DOU, Zhongsheng SUN, Yan TENG, Xiaoning LI
School of Mechanical Engineering, Nanjing University of Science and Technology
Xiaolingwei 200, Nanjing, Jiangsu Province, P. R. China(E-mail of corresponding author Zhonghua GUO: gzh@njust.edu.cn)

Abstract: In order to enhance the performance of soft actuator and overcome the problems existing in former soft
actuator, a new type of filament-polymer jamming actuator is proposed. The variable stiffness is realized by filamentand polymer combination and mix structure with jamming technology. By using such actuator, the large-scalevariable-stiffness gripper is proposed and the bending stiffness are measure through experiments. The variable
stiffness grippers successfully take the gripping operation while those grippers without variable stiffness fail to grasp
the movable objects. The results show that soft filament-polymer jamming actuator has better dynamic
characteristics and they are the better choice for fabricating the gripping ends.
Keywords: Pneumatic soft actuator, Fiber filament, Stiffness, Jamming

INTRODUCTION
Due to the advantages of soft material actuators, such as good deformation ability, unlimited freedoms, andadaptability to variable environments, it becomes an important element in automatic robots in recent years. As
the core components of soft robots, the performance of the soft actuators is critical for the whole performance ofthe robots. The deformation ability of the actuators has been extensively studied, such as the deformation of soft
actuators in artificial elephant trunk, jellyfish, and worms[1-5].
Stiffness is the important parameter of the soft actuator and it has the great influence on the dynamic
performance of the actuator and gripping ends. It is essential to study the stiffness characteristics of soft polymer
actuators. The difference in the material and structure of the soft actuator will result in changes in the stiffnessof actuator, which will complicate the actuator stiffness study.
In this study, new structure is proposed by using silicone and filament. Influences on bending stiffness from thematerial parameters are studied, such as diameter, coverage, inner wall thickness and air pressure of composite
fiber filament chamber. The results show that the bending stiffness increases with the increment of coverage
ratio and inner wall thickness. Although there is not monotonic function of the relationship between bending
stiffness and filament diameter, there is a maximum value in the experimental figure. The fiber material type has
little effect on the bending stiffness, and the bending stiffness increases with the increment of the air pressure.
Moreover, the soft structure is combined with jamming element[6], and the characteristics are increased
significantly during experiments. The proposed new actuator and gripper provides useful gripping ends forautomatic transferring devices especially dynamic transferring applications.

DESIGN AND FABRICATION OF SOFT FILAMENT-POLYMER ACTUATOR
The purpose of this study is to change the bending stiffness in the driving direction of a rectangular section
software actuator, adjusting the bending stiffness of the actuator by winding it with fiber filaments. To change
the fiber material type, filament diameter, cover ratio and composite layer position, a rectangular composite
actuator is designed, as shown in Figure 1.
The actuator is made of Ecoflex-0020 silica gel and filament. The filament refers to the polyester wax fiber(hard) or cotton fiber (soft). The total length of the actuator is 130mm, where one end is the fixed end of 15 mm,
and the other end is the sealed end of 10 mm. The middle part of 105 mm is the curved end of the actuator, thecross section is a rectangle of 10 mm by 10 mm, and the thickness of the wall is 5.5 mm. The composite
filament actuator is composed of a linear actuator, a fiber composite layer and an encapsulation layer. The
encapsulation layer is to improve the stability of the actuator. Considering that the fiber has a good adhesion to
the material of silicone, we use hard polyester waxy fiber and soft cotton fiber, diameter of separately 0.15 mm,

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0.3 mm, 0.55 mm, 0.80 mm and 1.00 mm; and the cover ratio is from 0 to 100%. The Inner wall thickness of theactuator is 2 mm, 3 mm and 4 mm.

Figure 1. Actuator with different parameters

As shown in Figure 2, the process for fabricating FRSA can be divided into three phases. At the first stage,assemble 3D printed molds(including the side end caps, the inner core, the casing and the upper cover). Then,pour ecoflex-0020 silica gel into a mold, which is cured at the room temperature. At the second stage, themanufactured inner core of the actuator is wound with filament. For effective winding of filaments, anexperimental device is designed by the authors to complete controllable filament coverage work. The device iscomposed of a winding part and an outlet part. The winding part is fixed on the test bench, and the outlet part is
fixed on a speed control rail. Combine the stepping motor to control the winding speed and the filament
coverage ratio is controlled according to the variable speed of the guide rail. At the third stage, the inner linear
filament is accomplished by another process of casting and packaging. The physical model when the package iscompleted.

Figure 2. Fabrication of a composite fiber filament soft actuator

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EXPERIMENTAL RESULT OF BENDING STIFFNESSIt is important to study the effect of composite fiber on the bending stiffness of actuators. However, it is difficult
to establish a mathematical model because of the non-linearity of soft materials.
The experimental data show that the bending stiffness increases with the increment of filament cover ratio.
However, when the cover ratio reaches the value of 85%, the bending stiffness tends to be stable. In addition,
the bending stiffness also increases with the inner wall thickness. Under the same cover ratio, the bending
stiffness of FRSA increases firstly and then decreases with the increment of the filament diameter. When the
diameter of the filament is 0.55mm, the bending stiffness gets the largest value.
Air
OutletVacuum
Pump

Air
Inlet
Air
Compressor
F

Jamming
ChamberForcesensorx

Displacement
sensorInflation
Chambers

Stepper motor
Motorized Stage

Figure 3. Bending stiffness test platform

C over ratio （% ）Figure 4. Bending stiffness with variable cover ratios and filament diameters

SOFT FILAMENT-POLYMER JAMMING ACTUATOR AND EXPERIMENT
By the former research, the soft filament-polymer actuator with the diameter of the filament 0.55mm has thelargest scale of bending stiffness. Using such parameters, we combine the polymer structure with jamming
element. Jamming element is a useful approach to change the actuator stiffness for soft element. As shown in
Figure 5, we made the soft filament-polymer and jamming element separately, and combined them together.
The fabricated gripper is made of three Filament-Polymer Jamming Actuators. We put them underwater and didsome experiments. As shown in Figure 6, we prepared a large transparent water tank. We put the gripper in thewater to grasp the object. After that, we moved the gripper inside the water and recorded the gripping operation.
Without variable stiffness, the grasping operation could be successful under static situation, while the object fell
off when the gripper was moved. However, the gripper with variable stiffness had better adaptation to dynamic
operations and the grasping operation was successful.

Figure 5. Soft Filament-Polymer Jamming Actuator

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Figure 6. Comparison between actuator with and without variable stiffness

CONCLUSION & FURTHER WORK
The variable stiffness grippers successfully take the gripping operation while those grippers without variable
stiffness fail to grasp the movable objects. The results show that soft filament-polymer jamming actuator has
better dynamic characteristics and they are the better choice for fabricating the gripping ends.
Our lab continue to study the enhancement of the stiffness and others characteristics of soft actuators andgrippers. We recently proposed a new type of bio-actuator and fabricated it successfully. It is inspired bytwinning plant. There are lots of twinning plant in the nature. As shown in Figure 7, the morning glory bends
around the branch as spiral motion and gets a good performance of gripping. We realized the artificial soft
actuator in our lab and found it a good way to increase the soft actuator characteristics. The gripping operations
are successfully for different objects shown in Figure 8. In future, we will get more new pneumatic soft
elements with better characteristics.

Figure 7. twinning plant

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The 11th JFPS International Symposium on Fluid Power 2020(a) ball above 200g(b) Hollow iron pipe(c) wire plier
Figure 8. Objects grasped by bio-gripper of our lab(d) disk

ACKNOWLEDGMENTS
The authors acknowledge Postgraduate Research & Practice Innovation Program of Jiangsu Province (NO.
KYCX20_0364), and Student Extracurricular Academic Research Fund Project of Nanjing University of
Science and Technology.

REFERENCES
1. POLYGERINOS P, WANG Z, OVERVELDE JT, et al. Modeling of soft fiber-reinforced bending actuators. IEEE
Transaction of Robot.31.778–789.
2. AGARWAL G, BESUCHEt N, AUDERGON B. Stretchable materials for robust soft actuators towards assistive
wearable devices.Sci Rep. 6.34224.
3. WANG L, IIDA F. Deformation in soft-matter robotics: A categorization and quantitative characterization. IEEE Robot
Automation Mag.22.125–139.
4. CHO K-J, KOH J-S, KIM S. Review of manufacturing processes for soft biomimetic robots. Int J Precis Eng Man.
10.171.
5. SUN Y, SONG S, LIANG X. A miniature soft robotic manipulator based on novel fabrication methods. IEEE Robot
Autom Lett.1.617–623.
6. GUO Z, LI X, SUN Z, et al. Soft shaping gripper inspired by marine animals. The 10th JFPS International Symposium on
Fluid Power Fukuoka 2017

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Robotic Blood Vessel Mechanism
for Self-Healing Function of Soft Robots

Kenjiro TADAKUMA*, Shohei INOMATA*, Yuta YAMAZAKI†,
Fumiya SHIGA†,Masanori KAMEOKA†, MD Nahin Islam SHIBLEE†,
Issei ONDA*, Tomoya TAKAHASHI*, Yu OZAWA*, Masahiro WATANABE*,
Hidemitsu FURUKAWA†,Masashi KONYO*, Satoshi TADOKORO** Department of Information Science Tohoku University, † Faculty of Science, Yamagata University
6-6-01 Azaaoba, Aramaki, Aobaku, Sendai Miyagi, 980-8579 Japan(E-mail: tadakuma@rm.is.tohoku.ac.jp)
Abstract. Soft robots, made of elastomers and gels tend to easily break, punctured, or torn by contact with sharp andthin objects. To deal with this problem, we had proposed a robotic blood vessel mechanism that has a self-healing
function by inserting liquid-filled tubes in the body. The liquid flows out and is cured which makes the would heal
when the body breaks. In this research, we have improved the healing performance by attaching a liquid-absorbent
material on the robot surface. Through experiment, it has confirmed that the self-healing ability can be improved byattaching the absorbent material.
Keywords: Mechanism, Self-Healing, Blood Vessel

1. INTRODUCTION
In recent years, research and development on soft robots have been actively promoted in the field of robotics.
Soft robots can be expected to be very useful from the viewpoint of application compared to conventional robots,as they can be softly contacted with the contact target and can be deformed flexibly according to theenvironment. However, many soft robots have their bodies composed of balloon shaped, membrane structures,and because of their flexible body structure, membranes are easily damaged by contact with the external
environment. This fatal issue such as breaking of bodies were still left.
On the other hand, various researches and developments have been promoted from the viewpoint of self-healingmaterials[1]-[7], and examples of sheets[8] or tires[9] that can be repaired with only small scratches haveappeared. However, there are still many studies at the conceptual stage, and self-healing itself takes several
hours, and self-healing is impossible if the wound is damaged at the level where the wound opens up[10]-[14].
In view of the above problems, we propose a basic concept of robot blood vessels as shown in the next section.
In this paper, the effect of the proposed basic principle was confirmed by explanation of the basic concept,
realization of the actual prototype models, and basic experiments were conducted with them.

Monomer solution
Reaction initiator(a) whole body healable robot with robotic blood vessels (b) co-axial type of robotic blood vessel
Figure 1. Concept of the Robotic Blood Vessels
1

Blood vessle

2

Normal Condition (Expansion)

Normal Condition

4

3

Wound
InjuredAfter Self-Healing

Figure 2. Blood Vessels Mechanism with Twisted Configuration
and Pouch Type Actuator Healable to Re-expanding by Robotic Blood Vessels

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2. BASIC CONCEPT
Fig.1 shows the basic concept of a robotic blood vessel. As shown in Fig. 1, blood vessels are stretched to every
corner of a soft robot with a flexible body, and even if a part is damaged, blood oozes from the wound, as if itwere platelets. The concept is to close the wound and repair it. As shown in Fig. 1 (b), in the cross section, aconcentric arrangement of a two-liquids mixture type liquid, a monomer solution and a reaction initiator, wasconsidered. However in actuality, because of the difficulty in connecting to the pipe joint in the above
concentric structure, in the first stage of the trial prototyping, the structure was adopted in which two pipes werearranged in a spiral shape as shown in Fig. 3. And table 1 shows the specification of the prototype model.

(a)(b)
Figure 3. (a) Prototype Model of the Blood Vessel Mechanism with Twisted Configuration, (b)Straight Typeand Twisted Type of the Blood Vessel Mechanism
Table 1. Specification of the First Prototype Model

3. BASIC EXPERIMENTS
3-1. Basic Mixing Experiment
As shown in Fig.4, two syringes containing two liquids stained in red and blue, respectively, were connected tothe tubes of the respective vascular mechanisms shown in figure. An experiment was conducted to observe themixing of two liquids.
First, when a cut was made in the direction perpendicular to the blood vessel, there was almost no difference
between the straight type and the helical twist type in the degree of mixing of the two liquids. Next, an
experiment was performed to observe the mixing of the two liquids when a wound was made in the direction
parallel to the axial direction with respect to the blood vessel(Fig. 5 and Fig. 6).
From the results of the experiments, it was confirmed that the proposed helical twisted type blood vessel
mechanism that was devised mixed more liquids through this experiment with the real prototype. For soft robots,
there is a method of arranging an infinite number of infinitesimal small blood vessels. However in reality, if thefinite number of blood vessels are arranged, this helical twisted type robot blood vessel devised is considered tobe very effective and useful in mixing the monomer solution and the reaction initiator for repairing themembrane structure.

Figure 4. Basic Experimental Setup

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1

2

3

14[s]

0[s]

19[s]

Figure 5. Experiment of mixing of two liquids when scratched in the axial direction (Straight Version)

2

1

3

66[s]

0[s]

102[s]

Figure 6. Experiment of mixing of two liquids when scratched in the axial direction (Twisted Version)
3-2. Basic Healing Function Experiment
Based on the above experimental results, assuming that the sheet on the surface of the pouch-shaped structure
shown on the right of figure 2 is damaged, one surface of the sealed box which take internal pressure was set toconfirm the basic healing function. The sheet with a built-in blood vessel developed this time was put at theupper part of the sealed box. The appearance of the experimental equipment is shown in the figure 7. The sameamount of each solution was secreted from two syringe pumps. In addition, in order to make it easier to create amixed state by keeping the exuded solution in one place, a cloth equivalent to gauze used in case of injury wasplaced on the surface of the sheet containing the blood vessels even if the direction of the gravity is changed
during healing process. If the surface of the sheet containing the blood vessels is injured, the two liquids that
correspond to the blood are mixed to close the wound on the sheet, and if it can withstand the internal pressureand expand, the repair will be successful.

Figure 7. Experimental equipment of Pressure Test

Figure 8. Self-Hearing Process of the Robotic Blood Vessel Mechanismand Expanding motion after healing without any leakage of the inner air

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Figure 8 shows the state of self-repair of the robot vascular mechanism with a cloth having a liquid-absorbing
function on the surface, and the state of expansion after self-repair by applying internal pressure without causing
air leakage.
As a result of the experiment, it was confirmed that the flexible sheet mounted on the upper surface of the sealed
box expanded without causing air leakage even when an internal pressure of 10 kPa was applied. From theabove, the basic self-repairing function of the devised robot blood vessel was confirmed by an experiment using
a prototype.

4. CONCLUSIONS AND FUTURE WORKSIn this paper, the basic concept of a robotic blood vessel for body repair of a soft robot was proposed. In
addition, in the helical twisted type vascular mechanism that was suggested, the effect of mixing in comparison
with the conventional method when the wound was made in the axial direction was confirmed through theexperiments with actual prototype models. In the future, considering the direction of gravity with respect to thewound, it is conceivable to add viscosity to the liquid and install a gauze structure on the surface, so that mucheffective healing can be performed without dripping the two liquids.

APPENDIX
The study of this robotic vascular mechanism is from the viewpoint of how to actively repair the body when it isdamaged. On the other hand, we are also conducting research and development with the cut resistance gripper
mechanism from the viewpoint of taking measures to prevent damage itself as much as possible by equipping itwith higher durability. In the durability performance test of this cut-resistant cloth against external contact, wehave devised a test device that basically emphasizes the method of "repeated contact at the same place". The tipof the cutting edge to be contacted is replaced with a sharp object (point contact plate, line contact plate) thatcan be regarded as a very small R, and the extent to which it can withstand different external contacts isinvestigated. In addition, the cloth tension and the contact position with the cloth can be adjusted by the tworollers on the left and right shown in the figure. The one shown in the figure A is a method with a constant angle,
but the one with a variable angle as the next step can be set as a stricter contact state from the viewpoint of cut
resistance.

Cut-Regsistant Clothfor Protection of Soft Robots
Left Roller

1

Touching Object with an Edge

Higer Tension

2

Cut-resistance ClothContact part
(Pointed shape)

Right Roller

Lower Tension

3

Roller for Adjust
the Tension of Cloths

Figure A. Durability Test Equipment for Contact Resistant ClothACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant in Aid for Scientific Research on Innovative Areas“Science of Soft Robot” project under Grant Number JP18H05471

REFERENCES
1. TERRYN, Seppe, et al. Self-healing soft pneumatic robots. science Robotics, 2017, 2-9: eaan4268.
2. WHITE, Scott R., et al. Autonomic healing of polymer composites. Nature, 2001, 409-6822, pp.794.

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3. CHEN, Xiangxu, et al. A thermally re-mendable cross-linked polymeric material. Science, 2002, 295-5560, pp.16981702.
4. CHENG, Nadia G., et al. Thermally tunable, self‐healing composites for soft robotic applications. MacromolecularMaterials and Engineering, 2014, 299-11, pp.1279-1284.
5. BILODEAU, R. Adam, KRAMER, Rebecca K. Self-healing and damage resilience for soft robotics: A review. Frontiersin Robotics and AI, 2017, 4, pp.48.
6. MARTINEZ, Ramses V., et al. Soft actuators and robots that are resistant to mechanical damage. Advanced Functional
Materials, 2014, 24.20: 3003-3010.
7. TERRYN, Seppe, et al. Development of a self-healing soft pneumatic actuator: A first concept. Bioinspiration &biomimetics, 2015, 10.4: 046007.
8. TERRYN, Seppe, et al. Toward self-healing actuators: A preliminary concept. IEEE Transactions on Robotics, 2016, 323, pp.736-743.
9. NAGAYA, K., et al. Tire with self-repairing mechanism. JSME International Journal Series C Mechanical Systems,
Machine Elements and Manufacturing, 2006, 49-2, pp.379-384.
10. TOOHEY, Kathleen S., et al. Self-healing materials with microvascular networks. Nature materials, 2007, 6-8, pp.581.
11. TOOHEY, Kathleen S., et al. Delivery of two‐part self‐healing chemistry via microvascular networks. Advanced
Functional Materials, 2009, 19.9, pp.1399-1405.
12. BLAISZIK, Benjamin J., et al. Self-healing polymers and composites. Annual Review of Materials Research, 2010, 40,
pp.179-211.
13. HANSEN, Christopher J., et al. Accelerated self ‐ healing via ternary interpenetrating microvascular networks.
Advanced Functional Materials, 2011, 21.22, pp.4320-4326.
14. DEAN, Leon M., et al. Enhanced Mixing of Microvascular Self-Healing Reagents Using Segmented Gas–Liquid Flow.
ACS applied materials & interfaces, 2018, 10.38, pp. 32659-32667.

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Genaral Session | Oil hydraulics

Robotics and Mechatronics 2[GS1-2-01] Design and Manipulability Analysis of a Redundant Anthropomorphic
Hydraulically Actuated Manipulator〇Fu Zhang1, Junhui Zhang1, Min Cheng2, Ruqi Ding3, Bing Xu1, Shen Zheng1 （1.State Key
Laboratory of Fluid Power and Mechatronic Systems, Zhejiang University, 2.State Key
Laboratory of Mechanical Transmissions, Chongqing University, 3.Key Laboratory ofConveyance and Equipment, Ministry of Education, East China Jiaotong University）[GS1-2-02] Wide Field of View Projection Display for Remote Control ofConstruction Robot〇Daisuke Kondo1 （1.Osaka University）[GS1-2-03] Experimental Implementation of a Hydraulic Turbine Access System with
Six-DoF Active Motion Compensation for Taiwan Offshore Wind Farms
Mao-Hsiung Chiang1, Bo-Yen Chen 1, 〇Sheng-Chia Lin1 （1.National Taiwan University）[GS1-2-04] Modular Hydraulic Servo Booster for Multi-Axis Robots
Sang-Ho Hyon1, 〇Tomoro Kai1 （1.Ritsumeikan University）[GS1-2-05] Comparison of Mechanical Drive System and Hydraulic Direct-Drive
System for Motor Power〇Juri Shimizu1,2, Takuya Otani2, Kenji Hashimoto3, Atsuo Takanishi2 （1.Hitachi, Ltd.,
2.Waseda University, 3.Meiji University）©The Japan Fluid Power System SocietyGS1-2-01

The 11th JFPS International Symposium on Fluid Power 2020

Design and Manipulability Analysis of a Redundant
Anthropomorphic Hydraulically Actuated Manipulator
Fu Zhang*, Junhui Zhang*, Min Cheng**, Ruqi Ding***, Bing Xu*, and Shen Zheng**State Key Laboratory of Fluid Power and Mechatronic Systems, Zhejiang University, China.
(E-mail: benzjh@zju.edu.cn)**
State Key Laboratory of Mechanical Transmissions, Chongqing University, China.
***
Key Laboratory of Conveyance and Equipment, Ministry of Education.
East China Jiaotong University, China.

Abstract. Heavy-duty rescue manipulator, replacing humans in dangerous working conditions, has the characteristicof higher robustness and a significantly larger power-to-weight ratio compared. The hydraulic driven robotic
manipulator, applied to the heavy-duty rescue manipulator, should have the characteristic of high capability of load,high accuracy, and high dexterity of action, while the existing hydraulic-driven robotic manipulators can hardly
satisfy because of the low system pressure, large moment inertia, the installed position of the control valve, andstructure. We develop a new type of spherical wrist joint to improve the flexibility of the robotic manipulator. And
we design a type of servo-controlled hydraulic-driven robotic manipulator to achieve heavy load and high control
accuracy. Then a performance index of the hydraulic manipulator is proposed based on constrained manipulability
polytope (CMP).
Keywords: Hydraulic manipulators, Manipulability

INTRODUCTION
Natural disasters and major emergencies seriously threaten the safety of human life and property, so the rapid
post-disaster emergency work and safe rescue have become an essential task in the field of public security inrecent years. However, the unstructured environment of the disaster scene makes it difficult for the rescue
workers to carry out the rescue work quickly, efficiently, and safely. Disaster rescue robots can replace humanbeings to carry out rescue tasks in a dangerous environment, with solid manipulability and high rescue
efficiency, which has been highly valued by all countries [1].
The large heavy-duty robot with rescue operation ability must have the capability of lifting, expanding, grasping,carrying, and other actions. Whether such actions are realized or not depends on the manipulator loading
capacity, action flexibility, and control accuracy. The hydraulically driven manipulator has the advantages ofhigh power-to-weight ratio and high robustness. It is suitable for the heavy load conditions of damaged roads,
building repair, collapse structure separation, etc. The heavy-duty robot with a hydraulic manipulator system
plays an irreplaceable role in the disaster scene [2].
In order to meet the needs of complex tasks in the rescue scene, the rescue manipulator must have the ability to
precise operation and good interaction with the operator. Following the configuration of the human arm, we
designed a complete hydraulic servo-controlled heavy-duty manipulator with integrated DOF joints [3-5].
Quantifying the manipulator's performance from the perspective of manipulability is especially important foranalyzing and designing new manipulator mechanisms and selecting appropriate manipulator configuration in
the grasping and operating environment. For the rescue manipulator system, the ability to monitor its motion is
crucial to the successful execution of the task. Therefore, the system must be equipped with methods to measure
its performance, accuracy, or success rate. In this case, the manipulability measurement is a common technique
to determine the manipulability of the workspace [6].
However, unlike an electric manipulator, due to the characteristics of a single-pump-with-multi-actuators system,
the velocities of different joints are coupled, and the joint velocity limits are variable. The velocity constraints ofvarious joints are determined by the area parameters of the hydraulic cylinder or the displacement of the swing
cylinder. Not only that, more precisely, velocity limits interact [7-9]. Therefore, it isn't easy to evaluate themanipulability of hydraulic manipulators by existing methods.
Yoshikawa analyzed the manipulability of the end effector of the manipulator and proposed an evaluation
method of the manipulability, an index related to the volume of the control ellipsoid [10]. Togai suggested using
the reciprocal of the condition number of the Jacobian matrix of the manipulator to evaluate the manipulability,
which represents the velocity isotropy of the end effector [11]. These local performance indicators cannot reflect
the overall performance of the manipulator. For example, the limit of joint position is not considered, and theaccessibility is not included in the evaluation. The global performance can also be evaluated by adding local

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performance indicators of the penalty function. The penalty function of the joint position is introduced into thescaling of manipulation ellipsoid. The ellipsoid axis near the joint limit is reduced so that the global position
accessibility index can be evaluated by local manipulation ellipsoid [12-13]. Abdel Malek et al. put the joint
position constraint into the augmented Jacobian matrix and optimized the placement of the manipulator by
analyzing the augmented matrix [14].
The manipulability ellipsoid cannot transform the velocity limits of the joint into task space, so it may not beable to give an accurate maneuverability evaluation index. Lee proposed scaling Jacobian matrix and steerable
polyhedron to consider the constraint of joint velocity. The former can't convert joint velocity into task spacevelocity accurately [15]. Lee proposed scaling Jacobian matrix and steerable polyhedron to consider theconstraint of joint velocity. The former can't convert joint velocity into task space velocity accurately. The
polyhedron-based analysis is more accurate [16-17]. The same conclusion is obtained in the study of parallel
robots [18-19]. Philip long proposed a new metric called constrained manipulability polyhedron (CMP), which
considers the kinematic structure of the system, including closed chain or composite sub mechanisms, joint
limits, and the existence of obstacles [20].
Due to the complex and significant velocity constraints of the hydraulic manipulator, it is difficult to evaluate
the manipulative performance of the manipulator by traditional ellipsoid analysis. Because of the velocity limitof the joint coupling of the hydraulic manipulator, it is very suitable to use constrained manipulability
polyhedron to analyze the operability. Therefore, it is very meaningful to propose a performance index byanalyzing CMP.
In this paper, we first introduce the rescue robot we designed, then present a new index to evaluate themanipulability of a hydraulic manipulator by using CMP and show an application example of a two degree of
freedom manipulator.

DESIGN OF THE MANIPULATOR
In this part, we first introduce a hydraulically actuated spherical wrist joint and then introduce a seven degree of
freedom hydraulic manipulator designed according to the configuration of a human arm. For a serial
manipulator, a separate actuator can drive each degree of freedom of the serial mechanism without couplingwith other degrees of freedom mechanisms or driving units, which not only provides the convenience of simple
control strategy and a small number of actuators but also brings the disadvantages of limited control accuracy,
low utilization of axial space of the whole machine and single local actionability. The advantages anddisadvantages of the parallel mechanism are opposite to those of serial mechanism. Considering the advantages
and disadvantages of the two mechanisms, the full hydraulic drive servo heavy-duty manipulator is proposed toadopt a hybrid scheme.

Design of the Spherical Wrist
Y
Revolute
joint

Yaw
ψ

Base
frame

Grasp
cylinderRollcylinderO

Universal
joint
Pitch
θ Z

φ
Cross
shaft
Yaw
cylinderPitchcylinder

Ulna

X

Roll

End
Grasp

RadiusHuman forearm

Figure 1. Structure of the roll-pitch-yaw spherical wrist.
The push-rod mechanism is considered a proper transmission way. Unlike electrically actuated areas, the use ofpure 3-DOF parallel wrist is limited in hydraulic areas since its size is too large to be installed on a serial

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manipulator, especially in compact environments. As shown in Fig. 1, Inspired by the human forearm that uses
radius and ulna to achieve wrist deviation, we use the hybrid serial-parallel concept to design the wrist as arevolute-universal chain, adopting two parallel linear cylinders to achieve pitch/yaw motion. The wrist structureis designed, as shown in Figs. 1 and 2.

D

ϕ

θ

C

ψ

A
B

Figure 2. Kinematic diagram of the roll-pitch-yaw spherical wrist.
Points A and B denote the kinematic pair centers of the pitch cylinder, while points C and D denote thekinematic pair centers of the yaw cylinder. The spherical RPY wrist is driven by two parallel linear cylinders forthe pitch/yaw motion and a rotary hydraulic cylinder for the roll motion. The pitch cylinder is connected to thebase frame and the cross shaft by two revolute pairs. The yaw cylinder is connected to the base frame by aspherical pair and a wrist plate via a universal joint. Moreover, the cross shaft is connected to the base frame via
a revolute pair to construct a closed-loop subchain with a four-bar linkage. Besides, the cross shaft is connectedto the wrist plate via a revolute pair. In Fig. 1, it should be noted that Point D is on the OZ axis, and Point B ison the YOZ plane at the zero position.

Design of the Overall Manipulator
Based on the proposed RPY wrist, a 7-DOF anthropomorphic redundant hydraulic manipulator is then designed.
Except for wrist pitch/yaw/roll, the other 4 DOFs for a human arm are shoulder pitch/yaw/roll and elbow pitch.
The 1st joint is determined as a revolute one. The motion of the 2nd joint is usually achieved by a linear
hydraulically driven component like the boom cylinder of the excavator’s arm. Thus, the basic configuration is
given in Fig. 3(a), in which the wrist joints are considered as the 5th, 6th, and 7th joints ( R5 , R6 , R7 ). Referringto the human arm layout, there are two selected configurations of the joints 2-4 ( R2 , R3 , R4 ) shown in Fig. 3(b)and (c). A typical spherical-revolute-spherical (SRS, R1 -R3 ) redundant structure is formed for the configurationin Fig. 3(b), which has been applied in electrically driven areas extensively, but resulting in that large
installation space and unnecessary axial length are difficult to avoid. Comparatively, the HRM shown in Fig. 3(c)is more like a human arm.
R4

R3
R2

R5 R6

R7

R4

R3
R2

R4

R3
R2

R1

Figure 3. Selected configurations of the 7-DOF hydraulic robotic manipulators
Through this configuration, the HRM and its three-dimensional model are designed as Fig. 4. It is pointed outthat the configuration of our designed hydraulic manipulator is different from that in Fig. 3 (c). In Figure 3 (c),the first three joints are perpendicular to each other and intersect, while the first three joints of our designed
hydraulic manipulator are perpendicular but not intersect, which is caused by the layout space of the hydrauliccylinder and swing cylinder.
The axis of the 3rd joint goes through the wrist center to easily conduct self-motion manifolds of the 4th joint. A
small offset between the 1st and 2nd joints is thereby allowed and assigned for layout consideration. The 1st,
3rd, and 7th joints are driven by hydraulic swing cylinders, and the others by hydraulic linear ones. A highspeed swing cylinder can drive the 1st joint with a gearbox due to its ample installation space.

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Figure 4. Configuration and three-dimensional model of the proposed hydraulic robotic manipulators

PERFORMANCE INDEX OF HYDRAULIC MANIPULATOR
Previously, we introduced the design of a 7-DOF humanoid hydraulic manipulator. The manipulability of themanipulator affects whether the task can be completed successfully, and it is an essential index of manipulator
evaluation. In this part, we introduce the classical manipulability ellipsoid and introduce how to use thereciprocal of the condition number to evaluate the manipulability of the manipulator in operation space. It’s
introduced how to construct the joint velocity polyhedron according to the pump and valve parameters. The joint
velocity polyhedron is constrained considering the joint position limits. The constrained manipulability
polyhedron is obtained through an affine transformation. An index is presented which is similar to the reciprocalof the condition number, and it is used to evaluate the manipulator maneuverability based on the constrained
manipulability polyhedron.

Manipulability Ellipsoid
The differential kinematics of an n-DOF manipulator can be written asv
=v n =
 J n qw 

(1)

where n is the degree number of freedom, v n is velocity vector in operation space, v is linear velocity vector
in operation space, w is linear velocity vector in operation space, J n ∈  6×n is the Jacobian matrix,q = [q 1 , q 2 , , q n ]T is the joint velocity vector.
In joint space, the unit hypersphere can be written as(2)q T q ≤ 1.
After affine transformation of the Jacobian matrix, the unit hypersphere of joint velocity becomes ellipsoid of
operation space. It can be given as=ε{v ( J J ) v ≤ 1}
T
n

n

T
n

−1

n

The manipulability index of the manipulator can be written as1γ=
cond(J )(3)(4)

where γ is a dimensionless quantity denoting the isotropy of velocity, cond(•) denotes condition number for • .

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Manipulability Polytope
A polytope,  V can be represented as the convex hull of its vertex set (V-representation), i.e.,
n
n


V
(5)
=
i 1, , n1=
 x: x= ∑ α i yi α i ≥ 0, ∑ α=i=i 1 =i 1
where yi is the i-th vertex of the polyhedron, and x is the point in the polyhedron. This polyhedral point set canbe written asH(6)=Ax ≤ bj×k
where A ∈  , j is the number of half spaces, k is the dimension of space.

In the form of  H , we can easily construct the joint velocity polyhedron according to the joint velocity range ofthe manipulator. It can be written as
I 
 q

(7)
 H  n  q ≤  max = −q min  −I n 
where I n is the n × n identity matrix, q max and q min are joint velocity limits.
The joint velocity polyhedron in the form of vertices can be defined by vertices as(8)V = {q 1v , q v2 ,..., q v2 n }where q iv is the i-th vertex of the joint velocity polyhedron.
The joint velocity polyhedron matrix can be written as q 1v   q1min q2min  qnmin
 min −1 qn
 v   min min max q 2   q1 q2  qnmin
−1 qn
(9)
=
Q =
        v   max max
max maxq 2 n   q1 q2  qn −1 qn 
An affine transformation transforms the joint velocity polyhedron, and the vertex of the polyhedron istransformed by the Jacobian matrix to get a new manipulability polyhedron vertex. The new polyhedron iscalled manipulability polyhedron. It can be written as(10)=v1v , , v v2 n } {J n v1v , , J n v v2 n } V {=

Velocity limits of the hydraulic manipulator
For hydraulic manipulator, the above results are extended. The hydraulic manipulator is a single-pump-multiactuator system, and the velocity limits of the actuator are affected by the direction of the velocity of theactuator. Here we use the simplification method to simplify the system. The single pump is divided into several
virtual pumps, and the flow of each pump is distributed in proportion to the rated flow of the generator. It can be
written as
Q(11)
Qi ,max = n i ,nom Qmax
Q
∑ i ,nom
i =1

where Qi ,nom is the norm flow rate of the i-th control valve, Qi ,max is the maximum flow rate of the i-th joint
control valve, Qmax is the maximum flow rate of the pump.
In the quasi-steady state, regardless of the pressure dynamics, the relationship between the joint angular velocity
and the flow rate of the hydraulic cylinder is followed as Qiηv
 V , i -th joint is driven by a swing cylinder or motor , i, g(12)qn = ∂qi Qiηv, i -th joint is driven by linear cylinder,∂xi S (qi ) A1,i , qi ≥ 0,
S (qn ) =  A2,i , qi < 0,

(13)

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Where ηv is the efficiency coefficient, Qi is the flow rate of the i-th joint, Vi , g is the displacement of the i-th
joint driver, A1,i is the area of the i-th joint cylinder without the rod, A2,i is the area of the i-th joint cylinderwith∂qi
is angular-linear coefficient of the i-th joint.
∂xi
The velocity limits can be determined by the following equations as Qi ,maxηv, i -th joint is driven by a swing cylinder or motor, Vi , gqi ,max =  ∂qi Qi,maxηv , i -th joint is driven by a linear cylinder, ∂x S (q )in
the rod, Qi ,maxηv, i -th joint is driven by a swing cylinder or motor ,−Vi , gqi ,min =  − ∂qi Qi ,maxηv , i -th joint is driven by a linear cylinder,∂xi S (qi )


(14)

(15)

By substituting Equation (14-15) into Equation (7), the joint velocity polyhedron of the hydraulic manipulator
can be obtained.

Positional Joint Limit Constraints
The joint velocity polyhedron obtained above does not consider the joint position limit, so it is necessary to
introduce the scaling factor of the joint limit penalty. The velocity polyhedron is constrained before convertingto the workspace. The influence of joint position limitation on maneuverability is considered. The scaling factor
can be written as max(qi , qi ) − qi 
(16)
ϕimax = 1 − qimax − qi min(qi , qi ) − qi (17)qimin − qiq min + qimax
(18)
q= i
2
Where ϕimax is the scaling factor of the upper boundary of the i-th joint velocity, ϕimin is the scaling factor of thelower boundary of the i-th joint velocity.

ϕimin = 1 − Constrained Manipulability Polytope
Equations (16-17) are introduced into Equation (7) to obtain the constrained joint velocity polyhedron. It is
given as φ max q max  In 

q
≤
 min

 −I 
 n −φ q min 

(19)

min
Where φ max= diag([ϕ1max , ϕ1max , , ϕnmax ]) , φ min= diag([ϕ1min , ϕmin
2 , , ϕn ]) .
The constrained joint velocity polyhedron is substituted into Equation (10) to obtain the constrained
manipulability polyhedron.
In the constrained manipulability polyhedron, an index similar to the reciprocal Jacobian condition number isused to evaluate the manipulator's manipulability. It is given as
w∗
(20)
µ= q
wq

Where wq∗ is the shortest distance from the end of a manipulator to the hyperplane of a constrained
manipulability polyhedron, wq is the longest distance from the end of the manipulator to the vertex of constraint
manipulability polyhedron.

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ILLUSTRATIVE 2-DOF PLANAR ROBOT CASE
Two hydraulic cylinders drive the joints of shoulder pitch and elbow pitch of the hydraulic manipulator. The
performance index of the proposed manipulator is illustrated in Fig 5. The kinematics model of the two degrees
of freedom of the hydraulic manipulator is established as
2
(21)
( L12
+ L222 − ( x2 + x02 ) 2 ) / 2 L12 L22 =cos(ζ 2 )(22)

2
+ L224 − ( x4 + x04 ) 2 ) / 2 L14 L24 =( L14cos(ζ 4 )

ζ= q + ζ 0

(23)

∂q2
=
− L12 L22 sin(q2 + ζ 02 ) / ( x2 + x02 )
∂x2

(24)

∂q4
=
− L14 L24 sin(q4 + ζ 04 ) / ( x4 + x04 )
∂x4

(25)

where L12 , L14 and L22 , L24 denote the link lengths, ζ 2 and ζ 4 are angles changing with the displacement of
the hydraulic cylinders, q = [q2 ,q4 ] is the joint angle vector, ζ = [ζ 2 , ζ 4 ] is the angle vector, ζ 0 = [ζ 02 , ζ 04 ] is theangle vector in the initial position, x2 and x4 are hydraulic cylinder displacements, x02 and x04 are lengths ofhydraulic cylinders in the initial position.

Figure 5. Shoulder pitch and elbow pitch.
In the table, we show the values of the parameters needed.
Parametera2

Table 1. The values of the parameters.

Value (Unit)

Parameter

Value (Unit)

ζ 04

A1,2

0.003117 (m2)A1,40.002155 (m2)0.87541 (m)ζ 02

1.5348 (rad)a4

1.061 (m)L120.8013 (m)0.6430 (rad)L14

0.6732 (m)L220.2000 (m)q2-0.7873-0.6090 (rad)L24

0.1500 (m)q4-0.2599-1.8345 (rad)The two degree-of-freedom mechanism is analyzed based on the manipulation ellipsoid and the manipulability
polyhedron.
The manipulability ellipsoids of the manipulator are shown in Fig 6(a). The manipulability analysis of themanipulator workspace is shown in Fig 6(b).
Based on manipulability ellipsoid, a hypersphere in joint space is transformed into the operation space throughthe Jacobian matrix. It can be considered that the 2-norm of joint velocity is less than 1. This focuses more on
the configuration of the manipulator than on the actual velocity limits of the joint. Using this method to analyze

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the operability of hydraulic manipulators will cause serious problems because the velocity limits of different
joints of the hydraulic manipulator are significantly different and coupled.

Figure 6. Manipulability analysis based on the ellipsoid.
The manipulability polyhedrons of the manipulator are shown in Fig 7(a). The manipulability analysis of themanipulator workspace is shown in Fig 7(b).
Based on manipulability polyhedron, a polyhedron in joint space is transformed into the operation space throughthe Jacobian matrix. It can be considered that the maximum norm of joint velocity is less than the velocity limit.
Velocity limits are considered directly. Therefore, the manipulability polyhedron can more accurately describe
the velocity characteristics of the end of the hydraulic manipulator.

Figure 7. Manipulability analysis based on the polyhedron.

DISCUSSION AND CONCLUSION
In this paper, a kind of hydraulic spherical wrist joint is designed and applied to the humanoid hydraulicmanipulator. Then, we review the evaluation methods of manipulator manipulability.
A method of calculating the velocity limits of the hydraulic manipulator is proposed. Manipulability polyhedron
analysis is developed, so a performance evaluation index for hydraulic manipulators is obtained. This index canevaluate the isotropy of the end of the manipulator. Finally, two joints of the manipulator are taken as examples
to compare the analysis results of the manipulability ellipsoid and the manipulability polyhedron. The results
show that the manipulability polyhedron is more suitable for the performance evaluation of the hydraulicmanipulator.
The manipulability polyhedron can also be applied to the performance analysis of the two arms, adding thepenalty function of the obstacle. Moreover, manipulation polyhedron analysis has excellent potential formechanism design and task learning of hydraulic manipulators, such as task learning, selection of hydraulic

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components, and linear hydraulic cylinder structure optimization (in Fig 5). This is a very potential analytical
tool. This paper only shows an example of a 2-DOF mechanism. In the future, we will analyze the designed 7DOF humanoid manipulator thoroughly.

ACKNOWLEDGMENTS
This work was supported in part by the National Natural Science Foundation of China under Grant 91748210,
the National Outstanding Youth Science Foundation of China (Grant No. 51922093), and Jiangxi Provincial
Excellent Young Funding Program (Grant No. 20192BCBL23005).

REFERENCES
1. HONG S, PARK G, LEE Y, et al., Development of a tele-operated rescue robot for a disaster response, International
Journal of Humanoid Robotics, 2018, 15-4, 1850008.
2. CHEN K, KAMEZAKI M, KATANO T, et al., Fundamental development of a virtual reality simulator for four-arm
disaster rescue robot OCTOPUS, 2016 IEEE international conference on advanced intelligent mechatronics (AIM). IEEE,
2016, pp.721-726.
3. KIVELÄ T, MATTILA J, AND PUURA J., A generic method to optimize a redundant serial robotic manipulator's
structure, Automation in Construction, 2017, 81, pp.172-179.
4. LIANG X, WAN Y, ZHANG C., Robust position control of hydraulic manipulators using time delay estimation andnonsingular fast terminal sliding mode, Proceedings of the Institution of Mechanical Engineers: Part I- Journal of
Systems and Control Engineering, 2018, 232-1, pp.50-61.
5. LUO X, LI L, XIE L, et al., Compliance Control Using Hydraulic Heavy-Duty Manipulator, IEEE Transactions on
Industrial Informatics, 2019, 15-2: pp.1193-1201.
6. JOHNS JIN L, LI S, LA H M, et al., Manipulability optimization of redundant manipulators using dynamic neuralnetworks, IEEE Transactions on Industrial Electronics, 2017, 64-6, pp.4710-4720.
7. LAMPINEN S, NIEMI J, MATTILA J., Flow-bounded trajectory-scaling algorithm for hydraulic robotic manipulators,
2020 IEEE/ASME International Conference on Advanced Intelligent Mechatronics (AIM). IEEE, 2020, pp.619-624.
8. HULTTINEN L, MATTILA J., Flow-limited path-following control of a double Ackermann steered hydraulic mobile
manipulator, 2020 IEEE/ASME International Conference on Advanced Intelligent Mechatronics (AIM). IEEE, 2020,
pp.625-630.
9. WANNER J, SAWODNY O., Tool-center-point control of a flexible link concrete pump with task space constraints
using quadratic programming, 2020 IEEE/ASME International Conference on Advanced Intelligent Mechatronics (AIM).
IEEE, 2020, pp.565-570.
10. YOSHIKAWA T., Manipulability of robotic mechanisms, The international journal of Robotics Research, 1985, 4-2,
pp.3-9.
11. TOGAI M., An application of the singular value decomposition to manipulability and sensitivity of industrial robots,
Siam journal on algebraic discrete methods, 1986, 7-2 pp.315-320.
12. CHAN T F, DUBEY R V., A weighted least-norm solution based scheme for avoiding joint limits for redundant joint
manipulators, IEEE Transactions on Robotics and Automation, 1995, 11-2, pp.286-292.
13. DARIUSH B, HAMMAM G B, ORIN D., Constrained resolved acceleration control for humanoids, 2010 IEEE/RSJ
International Conference on Intelligent Robots and Systems. IEEE, 2010, pp.710-717.
14. ABDEL-MALEK K, YU W, YANG J., Placement of robot manipulators to maximize dexterity, International Journal of
Robotics and Automation, 2004, 19-1, pp.6-14.
15. LEE J., A study on the manipulability measures for robot manipulators, Proceedings of the 1997 IEEE/RSJ International
Conference on Intelligent Robot and Systems. Innovative Robotics for Real-World Applications. IROS'97. IEEE, 1997,
3, pp.1458-1465.
16. SEKIGUCHI H, OHNISHI K., Force capability evaluation methods for bilateral controlled manipulators, 2017 IEEE
International Conference on Mechatronics (ICM), IEEE, 2017, pp.111-116.
17. FINOTELLO R, GRASSO T, ROSSI G, et al., Computation of kinetostatic performances of robot manipulators with
polytopes, 1998 IEEE International Conference on Robotics and Automation (Cat. No. 98CH36146). IEEE, 1998, 4,
pp.3241-3246.
18. KRUT S, PIERROT F., Velocity Performances Indexes for Parallel Mechanisms with Actuation Redundancy, Robotica,
2004, 22-2, pp.129-139.
19. MERLET J P., Jacobian, manipulability, condition number, and accuracy of parallel robots, 2006.
20. LONG P, PADIR T., Constrained manipulability for humanoid robots using velocity polytopes, International Journal of
Humanoid Robotics, 2020, 17-1, 1950037.

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Wide Field of View Projection Display for Remote Control ofConstruction Robot
Daisuke KONDO**

Dept. of Mechanical Engineering, School of Engineering, Osaka University
2-1 Yamadaoka, Suita, Osaka, 565-0871 Japan(E-mail: kondo@jrl.eng.osaka-u.ac.jp)

Abstract. A construction robot for disaster relief and its remote control platform is developed as a part of the
Impulsing Paradigm Challenge through Disruptive Technologies Program (ImPACT)’s Tough Robotics Challenge
Program. Unlike traditional remote-controlled construction equipment, the robot has two arms and many cameras.
For remote control the robot, a portable immersive projection screen is proposed. The screen which is in the shape ofa part of cylinder, has very simple and portable structure. The shape of screen derives the wide field of view. And it
allows to display various camera images in arbitrary arrangement.
Keywords: Rescue Robot, Remote Control, Immersive Projection Technology, ImPACTINTRODUCTION
The use of unmanned construction machinery will enable workers to avoid the dangerous conditions of theoccurrence of natural disasters. Under the Impulsing Paradigm Challenge through Disruptive Technologies
Program (ImPACT)’s Tough Robotics Challenge, an innovative construction robot for rescue was developed[1][2]. Fig. 1 shows the outer appearance of a construction robot.
The construction robot has dual arms for carrying out effective rescue activities in the disaster site and has manycameras. And it can be completely remote controlled from the teleoperation cockpit in distant site.

Figure 1. A construction robot.

The teleoperator of the construction robot visually identifies the surrounding situation to perform work
appropriately. The operations are performed based on the images transmitted from the many cameras settled on
the robot. In conventional remote-control system, the multi-monitor system is used for displaying multiple
images. However, the size and layout of images are fixed, so it is not comfortable to understand many images
for operator. In this paper, we propose a display designed for handling a number of images, and has panoramic
wide field of view.

THE DESIGN OF THE WIDE-FOV PROJECTION SCREEN
The teleoperation cockpit needs to be structed temporary in remote location near the disaster site. To facilitatetransportation and installation, we propose a method of forming a projection screen by elastically deforming aflexible plastic panel and fixing it to a frame. The structure is shown in Fig. 2. The projection screen was a 3mm-thick plastic plate. And it has a sand-blasted surface suitable for projection screen. The plate was elastically1

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deformed into a semi-cylindrical shape, and the upper both-side ends were fixed to the screen stand using
clamps and bolts. plastic guides were provided at nine points on the back, and metal stoppers were provided at
the bottom to further stabilize the shape. The restorative force of elastic deformation makes a seamless curved
surface. The semi-cylinder had a radius of 900 mm and a height of 1500 mm. To increase the downward view
angle, the semi-cylinder was inclined toward the back from the observer’s point of view. The angle was set as
26° to the vertical.
In general, most wide FOV screens have a concave spherical shape to give the user an immersive feeling. By
contrast, the proposed method involved a simple semi-cylindrical shape [3], which afforded a sufficient FOV
angle. The horizontal FOV angle at the assumed viewpoint was 133°. The semi-cylinder inclination reduced theless important upper FOV but provided a 56° downward field of view. When an immersive image is presented,
the ground conditions are viewed clearly. The viewing distance from the viewpoint to the screen was 1400 mm
in the front, which was sufficient.
Projector Frame

Projectors
Surface Mirror
Clamp

Guide

Stopper

Screen plate

Screen Frame
Figure 2. Image depicting assembly of screen and frames.

For image projection, four projectors were used. The projectors used were NEC NP-P451WJL. The screen
surface was partitioned into four areas (upper left, upper right, lower left, and lower right) and assigned to eachprojector. Front projection was used, and the projection was performed from above to avoid shadows caused by
the operator. To reduce the height of the system, the projected ray was folded using a surface mirror. Edge
blending was applied to the boundary of the area.
A PC is implemented for video processing. The PC can capture up to eight external video signals in HDMI
format or others and display them onto screen surface in real time. Generally, the image projected on a curved
surface is distorted, this distortion must be removed. Hence, free-form projection display technology 4] wasused to compensate the distortion. This technology generates images to be input to each projector by inverse
transforming the distortions caused by the relationship among the positions of the projectors, screen shape, andviewpoint position with very low latency. Finally, a unified correct image is displayed on the entire screen.

ARBITRARY ARRANGEMENT OF IMAGES
The various image data acquired by construction robot are transmitted to teleoperation cockpit. In conventional
multi-monitor system, because a number of images are displayed in fixed size and layout, it is not suitable tocheck the images. In our system, each screen is virtually arranged three-dimensionally in the space that can beseen through the field of view of the screen. They are compensated so that they appear as the correct plane andare oriented so that they are facing to the operator. Or it can be oriented at any angle. It is possible to arrange tooptimize usability by changing the size according to the importance.
The following eight channels of video were input to the display system.
(1) Front(Main View) (360 degrees camera)(2) Back View(360 degrees camera)(3) Arm camera(4) Gripper(5) Gripper (side)(6) Drone(7) AroundView(8) Robot’s posture2

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(1) to (7) were transmitted wirelessly from the construction robot, (8) is an image generated in local computer toindicate the robot’s joint angles and orientation in real time. The operator can choose some of these eight videos
as needed, and are displayed on the screen. the size and layout of images are switched arbitrary with a gamepad.
In addition, the arrangement according to the robot’s task can be stored in the system as a preset and can be
freely recalled and used. Some of the presets used are shown in fig.3. (a) is for traveling on rough terrain, (b) is
for manipulation using dual-arm and finding the survivor to rescue. The image necessary in each situation werechosen based on the opinions of the operator. Effective presets can be added according to the various work
purpose.

(6) Drone(1) Front(8) Posture(6) Drone(Main View)(Main View)(7) Aroundview(2) Back(8)Posture(1) Front(5) Gripper(side)(7) Aroundview(3) Arm(4)Gripper(a) Moving Mode(b) Manipulation Mode

Figure.3 Example of screen arrangement.

FIELD EXPERIMENT
Evaluation experiment was conducted in test field simulating a disaster site. The construction robot
demonstrates a scenario in which it was moving on the rough ground, after that find the survivor needs to help
under a collapsed house. As shown in fig.4, the proposed wide-FOV projection display was applied as the
remote-control cockpit for the evaluation experiment. The teleoperator completely performed remote control
using multi-screen.
The moving mode layout(fig.3(a)) was used during robot’s traveling, and manipulation mode layout(fig.3(b))
was used at situation that robot seeks a survivor. The main view (1) for checking the front of the robot is a wide
field of view image taken by the omnidirectional camera (Point Grey Research LADY BUG 5+) installed at thefront of the vehicle body. In order to reduce the amount of data transmission, images in the required directions
were cut out by image processing on the PC mounted on the robot, and the results were transmitted by wirelessvideo system(AMINMON CONNEX). The gripper view (4) rotates automatically on the screen to cancel theapparent rotation due to the angle of the robot’s gripper. The video delay on the main view is not small at 380ms.
The teleoperator can be adapted to the delay by practice. For further improvement of operability, this delay Thisdelay can be reduced by improving video transmission method or using the projector with a low delay mode.

Figure 4. Remote control cockpit.

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CONCLUSIONSAs a display system suitable for remote robot operation, we designed and implemented a simple and portableprojection-type wide-field display with a semi-cylindrical screen that is elastically deformed from a flat plate.
The display has portability to allow it to be installed in disaster areas in case of emergency. And it candisplay the multiple images which is captured by many cameras settled on the robot, at arbitrary positions.
Introducing this display into the remote-controlled cockpit of the construction robot platform in the ImPACT
Tough Robotics Challenge, and confirming that it can be used as a wide-field display as a wide-field display forremote-controlled operation.

ACKNOWLEDGMENTS
This study was supported by ImPACT Program “Tough Robotics Challenge” of the Council for Science,
Technology, and Innovation (Cabinet Office, Government of Japan).

REFERENCES
1.
2.

3.
4.

Yoshinada, H., Kurashiki, K., Kondo, D., et al., Dual-Arm Construction Robot with Remote-Control function, in
Tadokoro, S.(ed.), Disaster Robotics, Tracts in Advanced Robotics, vol. 128, Springer, 2019, pp.195-264.
Kurashiki, K., Yoshinada, H., Nagatani, K., Tanaka, M., Yamashita, A., Yokokohji, Y., and Konyo, M., A
Teleoperated Hydraulic Rescue Robot Integrated With Elemental Technologies, The 10th JFPS international
symposium on flued power, 1C12, 2017.
Kurokawa, M., Furukawa, M., Kondo, D., Ando, H., Maeda,T.,“Work efficiency for Remote Control with High Frame
Rate,” Proc. of the 2017 JSME conf. on Robotics and Mechatronics, 2A1-I08, 2017 (in Japanese).
Kondo , D. and Kijima, R., “Proposal of a Free Form Projection Display Using the Principle of Duality Rendering,”
Proceedings of Eighth International Conference on Virtual Systems and MultiMedia, pp.346-352,2002.

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Experimental Implementation of a Hydraulic Turbine Access Systemwith Six-DoF Active Motion Compensation for Taiwan Offshore
Wind Farms
Mao-Hsiung CHIANG1, Bo-Yen CHEN1, Sheng-Chia LIN1
1

Department of Engineering Science and Ocean Engineering, National Taiwan University
No.1, Sec. 4, Roosevelt Rd., Taipei 106, Taiwan
TEL: +886-2-3366 3730
FAX: +886-2-2392 9885
(E-mail: mhchiang@ntu.edu.tw)

Abstract. This study aims to investigate a new turbine access system (TAS) for offshore wind farms in the Taiwan
Strait under sea conditions with a wave period of 7.5sec and a significant wave height of 0.5m. The proposed system
includes the novel system integration of mechanism design and a hydraulic driving system. The dynamic cosimulation of TAS is achieved firstly to confirm the system design and parameters. The dynamic modelling of themechanism is implemented by ADAMS software. The hydraulic driving system and control system are derived
mathematically and implemented via MATLAB. Then, the dynamic simulation of TAS is achieved through the cosimulation of ADAMS and MATLAB to verify the active compensation control performance of TAS. After that, afull-scale test rig is set up for experimental verification. The vertical height, roll angle and vertical acceleration of the
TAS end effector can be reduced effectively through active motion compensation control, thus, improving the accesssafety of the offshore wind turbine.
Keywords: Offshore wind turbine access system, Active motion compensation, Motion compensation control,
Kinematics, Experiment.

INTRODUCTION
As the intensifying greenhouse effect and climate change, people began to pay attention to environmental
protection and renewable energy. In recent years, Taiwan achieved good results with rapid construction of
onshore wind farm, while its installation area has gradually become saturated. Relatively, typically out at sea,
there is a much higher wind speed allowing for more energy to be created at a time. Therefore, Taiwan isplanning to develop 5.5GW of offshore wind farm in Taiwan Strait by 2025. Maritime engineering has become
one of the important localization technologies in Taiwan. Assuring the safe maintenance of offshore windturbines has become a significant issue. This study aims to investigate the active compensation system suitablefor Taiwan Strait and applies to turbine access systems (TAS).
Maintenance staff enter the tower of offshore wind turbines 10 to 20 times per year. Assuring the safe
maintenance of offshore wind turbines has become a significant issue. In order to protect maintenance staff fromdangers when getting in and out offshore wind turbines, turbine access systems (TAS) are essential equipment.
TAS for offshore wind turbines not only ensure safety during maintenance, but also can effectively enhance theattendance of maritime ships and reduce the installation costs of maritime works.
In general, the docking methods for docking vessels at turbines primarily influence the design of TAS. Different
docking methods for connecting ship and wind turbines may lead to different design concepts for TAS. The
main categories [1] include:
I. pushing the vessel against the tower through propeller thrust,II. tying the vessel to the tower,
III. keeping the vessel at a short distance from the tower.
Docking methods (I) and (II) rely on the propeller thrust to maintain vessel stability. Docking method (III)requires a dynamic positioning system to determine vessel positioning. The additional equipment needs to beinstalled on the tower although docking methods (I) and (II) can effectively reduce the degree of freedom (DOF)
of vessel motion caused by waves. The TAS in docking method (III) can compensate six-DOFs vessel motion
caused by waves without additional equipment [2].
This study aims to investigate a new six-axial turbine access system for offshore wind farms in the Taiwan Strait,
under sea conditions with a wave period of 7.5 sec and a significant wave height of 0.5m. This study describes
novel system integration of the mechanism design, hydraulic driving system, and control system. The dynamic

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co-simulation of TAS is carried out firstly to confirm the system design and parameters through the cosimulation of ADAMS and MATLAB/SIMULINK. Through this co-simulation, we verify the effect of theactive compensation control system for TAS. Next, a full-scale test rig is set up for experimental verification.
Through the active compensation control experiment in the full-scale test rig, the 6-DOF motion of the TAS end
effector can be reduced and verified by practical experiments.

LAYOUT OF TAS
For the docking method that keeps the vessel at a short distance from the tower, this study applies Stewart
platform to TAS to compensate the motion caused by waves in six degrees of freedom. It consists of six
extensible actuators. Each actuator is connected to the moving platform by a set of joints consist of a universal
joint and a revolution joint. And the base platform is connected by a universal joint. The linear extension andretraction of the six actuators gives the platform six degree of freedoms positioning capabilities, including threetranslational and three rotational degrees of freedoms. The actuators labelled by 1 to 6 with anticlockwise, as
shown in Fig. 1. The length of the actuator by zero extension is 1.3 m and the length of maximum elongation is
1 m. The material of the platform is A36 steel which is a standard steel alloy that is a common structural steel inmechanical industry. It has density of 7800 kg/m3 and yield strength of 252 MPa and an allowable bending
stress of 154 MPa. The properties of ASTM A36 steel allow it to deform steadily as stress is increased beyond
its yield strength.
The system configuration of the six-DOFs TAS test rig is shown in Fig. 2. The position sensors are set inhydraulic cylinders to measure the cylinder position. In the control system, the D/A interface card sends out thecontrol signal to drive the hydraulic servo valve, and the A/D interface card receives the feedback signals fromposition sensors. Moreover, a hydraulic power unit with supply pressure of 70 bar is set up.

Figure 1 ADAMS model of TAS

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1

A/D
3

2

D/A

4

Figure 2 System configuration of TAS test rig

KINEMATICS OF TASThis section introduces the kinematics of TAS, including inverse kinematics and forward kinematics. Here, boththe inverse and forward kinematics of TAS are solved by MATLAB.
The inverse kinematic transformation for the TAS determines the required actuator lengths for a given
configuration composed of Cartesian position and orientation of the moving platform with respect to the baseplatform. As shown in Fig. 3, the origins P and O are the corresponding geometric center of the moving
platform and the base one. The Cartesian coordinates of the moving platform are given by the position of origin
P with respect to the fixed frame, designated by [ px py pz ]T , and the orientation of the moving platform isusually represented by three standard Euler angles [  , ]T , where px , py , pz represent the longitudinal(surge), lateral (sway), and vertical (heave) motions, respectively, whereas ,  represent the roll, pitch, andyaw motions. Each leg i links the attachment points Pi and Bi on the moving platform and the base, respectively.
In this configuration, the attachment points Pi and Bi can be expressed as [3] rP sin(Pi )   Pix   Pi   rP cos(Pi )    Piy  ,  P 0 iz i P

3 2
Pi  P (i 1) P

 rB sin(Bi )   Bix   Bi   rB cos(Bi )    Biy  ,  B 0 iz 

i  B
3 2
Bi  B (i 1)   B

Pi 

Bi 

i  1,3,5(1)

i  2, 4, 6

i  1,3,5(2)

i  2, 4, 6

where rP and rB denote the length between attachment points and the center of the moving and the baseplatforms, respectively. As shown in Fig. 3, θ and ϕ are the angles from the center of platforms to the adjacentattachment points on the moving and the base platforms, respectively.
The inverse kinematic problem of the TAS is to determine six link-lengths of the manipulator for particular
position ( px , py , pz ) and orientation (  ,  ) of the moving platform of the manipulator. Given a set of
position and orientation of the moving platform, the necessary link-vectors, denoted by li ( i = 1,2,…,5,6), canbe straightforwardly computed by using the following formulal i  R  Pi  D  B i

(3)

where

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 CR  R R  R    0
  SS   1 0
0  C  S
1 0  0 C  S    S C
0 C  0 S  C   0
0


0

0 

0 
1 

 r11 r12 r13  r21 r22 r23  r31 r32 r33 (4)is the rotation matrix of the platform and D=[ px py pz ]T which is the translation vector of the platform. Pi and
Bi are the translational offset vectors of link i on the moving platform and the base respectively. S and Caredefined as sin() and con(). The relation between the vectors can be seen in Fig.3.
The norm of li , i.e., || li ||, represents six link-lengths of the current pose, using Eqs. (1), (2) and (4). Thus, Eq. (3)can be written asli

2 px 2  p y 2  pz 2  rP 2  rB 2  2(r11 Pix  r12 Piy )( p x  Bix )2(r21 Pix  r22 Piy )( p y  Biy )  2(r31 Pix  r32 Piy ) pz  2( p x Bix  p y Biy )(5)

for i  1, 2,...,5, 6

As shown in Eq. (5), for a particular pose of the moving platform there will be an unique solutionli ( i = 1,2,…,5,6), namely, six link-lengths, thus the inverse kinematic problem of the TAS is nearlystraightforward.
When six-DOFs ship motion caused by waves are determined as roll, pitch, yaw, Zsurge, Zsway and Zheave, themotion of the base platform of TAS in 6-DOF compensation broll, bpitch, byaw, Zbsurge, Zbsway and Zbheave, can becalculated by DH method [4]. In order to let the moving platform stable, the particular position ( px , py , pz ) andorientation (  ,  ) of the moving platform of TAS in 6-DOF compensation must be determined aspx   Z bsurge ,p y   Z bsway ,pz   Z bheave  broll ,  bpitch ,  byaw(6)

The forward kinematic problem is a very difficult and challenging one that has initiated much researchattentions during the past decade. The problem is to determine the position ( px , py , pz ) and orientation(  ,  ) of the moving platform from six given actuator lengths li (i = 1,2,…,5,6), it is difficult to solve theproblem because of the inherent nonlinearity and complexity. The general form of the problem can be obtained
from Eq. (5) as follows:px 2  p y 2  pz 2  rP 2  rB 2  2(r11 Pix  r12 Piy )( px  Bix )2(r21 Pix  r22 Piy )( p y  Biy )  2(r31 Pix  r32 Piy ) pz  2( px Bix  p y Biy )  l i

2

0

(7)

for i  1, 2,...,5, 6

Figure 3 Geometry of moving and base platform and vector representation of TAS

Eq. (7) is multivariate polynomial equations set with respect to the moving platform position parameters( px , py , pz ) and trigonometric functions of orientation parameters (  ,  ). Therefore, the solution to theforward kinematic problem involves the solution of a system of highly nonlinear coupled algebraic equations invariables describing the moving platform pose. The Newton-Raphson method is often used to solve the problem,
but the duplicate steps before solution convergence cause the real-time application inaccessible. Furthermore,such method may cause infinite loops provided selection of wrong initial values.

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RESULTS AND DISCUSSION
This section contains an experiment of the test rig to verify the TAS control performance. To compensate for thewave-induced motion of the end effector, the required displacement of each hydraulic cylinder is determined by
the inverse kinematics, which become the target position of each hydraulic cylinder for the closed-loop position
control of the hydraulic system. Furthermore, the PID algorithm is used to design the controllers making thehydraulic cylinders generate force correspondingly to drive the TAS.
Figure 4 shows the block diagram of the overall TAS control system in the experiment. The control input fromthe controllers designed by PID is given to the hydraulic systems to drive the test rig, and the position of eachhydraulic cylinder is measured by position sensors and fed back to the controllers. Furthermore, based on themeasured position of the hydraulic cylinders, the position of the TAS end effector can be calculated according to
the forward kinematics.

Sea
Condition

Roll Angleof Ship
Motion+Pitch Angleof Ship
Motion+-+Yaw Angleof Ship
MotionSurgeof Ship

InverseKinematics+-

Cylinder 1
PID
Controller

X6T1

Cylinder 2
PID
Controller

X6T2

Cylinder 3
PID
Controller

X6T3

Cylinder 4
PID
Controller

X6T4

Cylinder 5
PID
Controller

X6T5

Sway
Displacementof Endeffector

Cylinder 6
PID
Controller

X6T6

Heave
Displacementof Endeffector

MotionSwayof Ship+-

MotionHeaveof Ship+-

MotionRoll Angle ofEnd effector

Pitch Angle ofEnd effector

Yaw Angle ofEnd effector

ForwardKinematics

Surge
Displacementof Endeffector

VerticalAccelerationof Endeffector

Figure 4 Block diagram of overall TAS control system in experiment

The direction relation between the ship and the wave is shown in Fig.5. Since the sea wave condition influences
the ship motion, the motion variation of the ship’s center of gravity caused by the given sea wave conditions of
the offshore wind farm in Taiwan, with a wave period of 7.5 sec, significant wave height of 0.5m and wave
direction of 45 degree is shown in Fig. 6. The wave data in this study is provided by Ship and Ocean Industries
R&D Center. Furthermore, the PID controller are designed by Ziegler and Nichols method.
According to the ship motion in Fig. 6, the desired position of the six hydraulic cylinders to compensate for the
motion of the end effector (the center of moving platform) can be calculated by the inverse kinematics, which
are given as the target position trajectory of each cylinder. Figure 7 shows the position control responses andcontrol errors of the hydraulic cylinders in the simulation and experiment, where the control errors of the six
hydraulic cylinders are all within 0.08m in the simulation and 0.11m in the experiment.
Furthermore, the active motion compensation response at the TAS end effector can be compared and confirmed
through the motion of six-DOF of the end effector. Figures 8 and 9 show the active motion compensation results
at the TAS end effector, including three displacement responses and three angle responses of the end effector. In
the condition without compensation, the roll, pitch and yaw angle significantly change within 3.6, 3.3 and1.2, respectively; the surge, sway and heave displacement significantly change within 0.23m, 0.23m and0.36m, respectively. After the active motion compensation, the surge, sway and heave displacement and theroll, pitch and yaw angle and the vertical acceleration change of the end effector can be respectively reduced to
1.4, 1, 0.39, 0.04m, 0.04m, 0.1m in simulation and can be respectively reduced to 2, 1.5, 0.5, 0.05m,
0.05m, 0.12m, in experiment. Therefore, through the verification and comparison in the simulation andexperiments, both controls can achieve motion compensation effects.

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Figure 5 Direction of wave

Figure 6 Ship motion caused by sea wave of 7.5 sec period, significant wave height 0.5m and wave direction 45 degree(a) Pitch angle, (b) Roll angle, (c) Yaw angle, (d) Sway position, (e) Surge position, (f) Heave position

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Figure 7 Hydraulic cylinder position control response and control errors(a) 1st hydraulic cylinder position response, (b) 2nd hydraulic cylinder position response,(c) 3rd hydraulic cylinder position response, (d) 4th hydraulic cylinder position response,(e) 5th hydraulic cylinder position response, (f) 6th hydraulic cylinder position response

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Figure 8 The motion compensation angle responseat end effector(a) Pitch angle response of end effector,(b) Roll angle response of end effector,(c) Yaw angle response of end effector

Figure 9 The motion compensation displacement
response at end effector(a) Surge displacement response of end effector,(b) Sway displacement response of end effector,(c) Heave displacement response of end effector

CONCLUSIONSThis study developed a TAS with a 6-DOF active motion compensation system for offshore wind farms in the
Taiwan Strait, under sea conditions with a wave period of 7.5 sec and significant wave height of 0.5m. The
system is driven by active motion compensation control systems using PID. The dynamic co-simulation of TAS
via ADAMS and MATLAB/SIMULINK were achieved to confirm the system design and parameters. In
addition, a full-scale test rig combined with the Real-time Workshop of MATLAB/SIMULINK was set up forexperimental verification. Under the sea conditions of the Taiwan Strait with a wave period of 7.5 sec andsignificant wave height of 0.5m, the motion in 6-DOF of the TAS end effector can be reduced effectively
through active compensation control in both the simulation and experiments. Finally, the TAS proposed in thisstudy has been verified in simulation and experiments with satisfactory active motion compensation
performance.

ACKNOWLEDGMENTS
This research was sponsored in part by Ministry of Science and Technology Taiwan under the grant MOST
MOST 110-3116-F-002-001.

REFERENCES
1. Twidell, J. and Gaudiosi, G., "Offshore Wind Power", 1 ed., Multi-Science Publishing Co. Ltd., 2014.
2. Cerda Salzmann, D. J., "Development of the Access System for Offshore Wind Turbines," Ph.D. dissertation, Delft
University of Technology, 2007.
3. Harib, K. and Srinivasan, K., "Kinematic and dynamic analysis of Stewart platform-based machine tool structures,"
Robotica, vol. 21, pp. 541-554, 2003.
4. Hartenberg, R. S. and Denavit, J. "Kinematic synthesis of linkages [M]." New York: McGraw-Hill 1964.

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Modular Hydraulic Servo Booster for Multi-Axis Robots
Sang-Ho HYON* and Tomoro KAI*
Department of Robotics, Ritsumeikan University
1-1-1 Noji Higashi, Kusatsu, Shiga, 525-8577 Japan(E-mail: sangho@ieee.org)*

Abstract. A new modular hydraulic circuit suitable for multi-axis robots is proposed. The circuit is characterized bymodular design, pressure boost, flow summing, and high response. The main module, called Boost Servo Pump
Module (BSPM), is composed of a servo pump and switching valves. The circuit can be simply plugged into
conventional servovalve systems, and improves the energy efficiency significantly. The paper presents the realizationof the circuit and possible operating modes, then presents the control method for each mode. The simulation results on
position control of two-link serial manipulator validate the proposed method by its clear improvement of tracking
performance especially in the shared boost mode, where small pumps work together to generate high pressure andflow to the actuator.
Keywords: Modular, Hydraulics, Robot control, Flow summing, Pressure boost

BACKGROUND
To date, most high-performance multi-axis hydraulic manipulators have been equipped with servovalves, which
are great inventions, nevertheless present shortcomings related to cost and energy efficiency. In this conventional
method, the supply pressure must be set to the highest pressure for the axis of the highest load, which causes
throttle loss at other axes. Electric hydrostatic actuators (EHAs) can solve this problem completely [1]. The
method is called volume control as opposed to throttle control. Each actuator provides fluid power by itself, whichallows decentralized power management. However, the servomotor must be larger as the required power becomes
higher. This can pose a severe obstacle against EHA applied to multi-axis robot control.
In 2013, the authors invented a new hydraulic circuit, referred to as the Hydraulic Hybrid Servo Booster (H2SB)[2]. As shown in Fig. 1, the circuit has a main pump, four valves (V1 to V4), and a servomotor-controlled pump(SP), which are connected to the inlet/outlet of the cylinder. It embeds a small SP into the valve bridge so that thehigh-speed (up to valve size) and precision (up to the resolution of SP) are achieved in a hybrid and cost-effective
manner. However, this method had three problems:(P1) Low dynamic response at the “open-circuit mode” due to low-cost cartridge valves(P2) Increased system size due to many independent valves,(P3) Low speed at the “closed-circuit mode” and the “boost mode” due to the small volume of SP

Common return line
V3

V4

J1
V1

M

J2

JN

SM

SM

V2

SM

Common pressure line
Figure 1. (Previous) Hydraulic Hybrid Servo Booster (H2SB) [2] for multi-axis robot. If the SPs are removed, the circuit
just works as the independent metering valve system [3,4,5] by controlling four proportional valves V1 to V4. When theoutlet valves, V3 and V4, are closed, then it works as independent EHAs with pre-loaded pressure. Moreover, SPs can boost
the pressure to obtain higher force and high-resolution flow independently by opening the diagonal set of valves.

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This experience led us to invent a new modular hydraulic servo system suitable for multi-axis robots in thispaper, where small servovalves and small servo-pumps are both utilized. By doing so, we can keep the highresponse of the servovalves while reducing the throttling loss at the valves as much as possible even for the lowpressure operation. Instead, we have abandoned the closed-circuit configuration.
The new circuit is designed to have a modular structure that meets three conditions: 1) the modules are one-toone connected to the actuators of the robot joints by hoses; 2) servovalves are directly mounted to the actuators;3) modules are connected to a common pressure source (pump) and reservoir. Thus we solve (P1).
Then, how about (P2) and (P3)? This question is related to what kind of circuit should be implemented in themodule. This led us to an idea of shared boosting, as opposed to decoupled boosting (already achieved in theprevious circuit in Fig. 1).

PROPOSED MODULAR HYDRAULIC CIRCUIT
Overview
Figure 2 shows the hydraulic circuit first proposed in this paper. The system has servovalves, two kinds ofpumps (the main pump and SPs), and switching valves, which are organized in modular configuration. The main
module is called Boost Servo Pump Module (BSPM), composed of a servo pump and a switching valve (normallyopened, three-positions, four-ports valve in this specific realization). There is also a special module called PowerTake-Off Module (PTOM). The BSPMs work in the following five operation modes, while PTOMs operate ateither N-mode or PTO-mode.

Figure 2. (New) Modular Hydraulic Servo Booster. The circuit is stacked together in one unit, and connected by hoses tothe actuators, to which the servovalves are directly mounted. The return lines may be combined to a single hose. Each BSPM
has a servo pump and a switching valve, but PTOM does not have SP because high-pressure flow may be provided by some
BSPMs. Basically, the number of modules is assumed to be N, which is the same as the number of the actuators, but the
numbers of BSPMs and PTOMs can be determined according to the motion tasks of the robot. Note that he robot does notneed to be a serial manipulator. Each module has P (common pressure), T (tank), S (shared) rails and Outlet/Return ports.
When the valves are OFF state, all the outlet of the SPs and the inlet of the servovalves in this circuit are connected throughthe S rail (called shared boost mode). This figure shows a specific realization employing commercial four-ports cartridge
valves, but one can replace them with multiple two-ports valves or custom-made three-ports valves. Without SPs andswitching valves, the circuit merely serves as a manifold for conventional servovalve systems (upper compatible). Note thatthe main pump could be powered-off because each SP can directly sack the oil from the reservoir through the line connecting
P and T through the check valve.

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Operation Modes
1) At low load conditions, the servovalves ensure a high response within the common supply pressure P0 andmaximum flow rate of the valves. Let us call this normal mode (N-mode). This mode can be achieved bycommanding +1 to the switching valves (right position). The main pump controls the common pressure tosome value 𝑃𝑃0 , for example, to keep some stationary posture of robots.
2) At higher load conditions, each SP boosts the pressure locally. This is called decoupled boost mode (DBmode), where we can set independent pressures, 𝑃𝑃1 , 𝑃𝑃2 , … , for each module. This mode also can be achievedby commanding +1 to the switching valves (right position).
3) Some SPs can share the output flow by summing, where the pressure is determined to the common pressurelarger than 𝑃𝑃0 . This is called shared boost mode (SB-mode). This mode can be achieved by commanding 0
to the switching valves (normal position).
4) Some SPs “devote” their flow to other modules by commanding -1 to the switching valves (left position).
This is called assisted boost mode (AB-mode). This mode is useful to have higher flow to specific actuatorgroups.
5) The last mode is trivial, power-take-off mode (PTO-mode). It just stops the SP (e.g. for cooling purpose)and take flow from other modules through the S rail.
[Note] It is impossible to achieve two different pressures in SB mode. In that case, we just separate the modules
into two groups, then block the S ports between the group (e.g. by inserting plugs).
One can operate the circuit in aforementioned five modes according to the loading conditions of robots. One
drawback of the circuit is that the meter-out control for the negative load (second and fourth quadrants) causes
the energy loss by throttling. This is the expense of holding dynamic response of servovalves for high-performancerobots. Therefore, it is important to set the common pressure as low as possible for some given robotic tasks. In
this case, the noise and temperature can be reduced significantly by using low-pressure quiet pumps (commerciallyavailable) for the main pump.
In our previous local report [6], we have presented a single-axis version of this circuit and verified the reductionof energy loss compared with conventional servovalve systems without scarifying the accuracy. But the previous
circuit had no flow routing valves, nor the modular structure. When we apply this circuit to serial manipulators,
where the volume of the proximal actuators is usually larger than the distal ones, we need to use larger pump andservo motors to the proximal joints. Mixture of pumps and motors of different sizes prevents the system frombeing compact and modular. The solution to this problem is to introduce the switching valves to transfer the flows
from the distal pump to proximal actuators (or vice-versa) as necessary. This report presents the modified circuitand simulation results.

Related Study
The idea of using multiple power sources to drive multiple actuators is of course not new. Theoretically, one canuse many On/Off valves to distribute the flow from multiple power sources to realize desired pressure/flow of theactuators. This is the very attractive feature of the hydraulic drive systems that the electric counterparts do not
have. There are many possible combinations, but one of the extremes would be digital hydraulics, which isextensively surveyed in [7]. Although the impressive invention would be the novel components such as digitalvalves, there are still many variants relying on conventional (low-cost) components such as cartridge valves. For
example, in [8], the authors proposed N-by-N (conventional) On/Off valves and (conventional) proportional
valves for N-DoF excavator. However, the system size and complexity of the controller are increasing as the DoF
increases. In [9], the authors combined an EHA with meter-out valve, then applied to 2-DoF robot motion control,
wherein the throttle loss is reduced by transferring the flow from one actuator to other actuator. In [10], aninteresting example of an excavator is shown, where the flow from bucket and arm cylinders is transferred to theboom cylinder to increase the speed.
All these studies are related to our work, but our circuit in Fig.2 uses servovalves and boosted pressure suppliedby multiple servo pumps, which are stacked in a modular configuration. Also, it is ensured that no actuators areblocked because the N-mode is always available. The authors think these points are unique, and attractive forhigh-performance dynamic hydraulic robots such as humanoid robots, whose joints are driven by many
servovalves [11].

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CONTROL FOR TWO-LINK MANIPULATORThis section presents an illustration of the basic control scheme for the proposed circuit. For simplicity, we show
the N=2 case and explain how the circuit is utilized to a 2-DoF serial robot shown in Fig.3, left. Specifically, we
restrict ourselves to a vertical lifting motion of the payload, where five phases exist as indicated in Fig.3, right.
Details of the modes and control laws, according to the temporal phase transition of the robot motion, areexplained as follows.
P hase1 P hase2 P hase3 P hase4 P hase5

0.8

X

0.6
0.4

X

Z

0.2

J1

0
-0.2

+θ1
0.4 m

J2

-0.4

Z
0

0.45 m1

2

3

Time [s]Joint 1 State

-θ2

-75

60

-80

40

-85

20
0

5

Joint 2 State

-70

80

8 kg

4

-90
0

1

2

3

Time [s]

4

5

-95

0

1

2

Figure 3. A two-link manipulator and its target trajectory of the tip position

3

Time [s]45Phase 1 (0~1 [s]):The both actuators operate at DB-mode because the high pressure is required for both joints. Instead of merely
generating high-pressure by boosting, each SP and servovalves coordinately control the position of cylinder toreduce the throttling loss. That is, SP generates the necessary inlet flow to achieve the desired velocity of thecylinder, while the servovalve control the outlet flow by using the actual (measured) pressure drop (meter-out
control). This is same as the classical load sensing hydraulic systems.
Let 𝑣𝑣̅ be the desired speed of the cylinder, which can be obtained by differentiating the desired position 𝑦𝑦�. The����������
target value for the inlet flow 𝑄𝑄𝚤𝚤𝚤𝚤 and outlet flow 𝑄𝑄𝑜𝑜𝑜𝑜𝑜𝑜 are computed by
𝐴𝐴𝑎𝑎 𝑣𝑣̅ ,𝑣𝑣̅ ≥ 0����
(1)
𝑄𝑄𝚤𝚤𝚤𝚤 = �𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝐴𝐴𝑏𝑏 𝑣𝑣̅ ,𝐴𝐴𝑏𝑏 𝑣𝑣̅ ,𝑣𝑣̅ ≥ 0������
(2)
𝑄𝑄
𝑜𝑜𝑜𝑜𝑜𝑜 = �𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝐴𝐴𝑎𝑎 𝑣𝑣̅ ,with the piston area of the cap-side 𝐴𝐴𝑎𝑎 and rod-side 𝐴𝐴𝑏𝑏 . Therefore, the commanded speed of the SP����𝜔𝜔2 should be����𝑄𝑄𝚤𝚤𝚤𝚤
(3)𝜔𝜔�=𝐷𝐷𝑝𝑝 ∙ 𝜂𝜂
where 𝐷𝐷𝑝𝑝 and 𝜂𝜂 are the volume and efficiency of the SP, respectively. For position tracking purpose,
we modify this by����
𝑄𝑄𝚤𝚤𝚤𝚤 + ∆𝑄𝑄𝑝𝑝
(4)
𝜔𝜔
�=
𝐷𝐷𝑝𝑝 ∙ 𝜂𝜂
where𝑣𝑣 ≥ 0
𝐾𝐾𝑝𝑝 (𝑦𝑦� − 𝑦𝑦)(5)∆𝑄𝑄𝑝𝑝 = �(𝑦𝑦
−𝐾𝐾𝑝𝑝 � − 𝑦𝑦) 𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒is the feedback term with the proportional gain 𝐾𝐾𝑝𝑝 .
The commanded input to servovalves are given by
����
𝑄𝑄𝚤𝚤𝚤𝚤 + ∆𝑄𝑄𝑣𝑣𝑣𝑣𝑣𝑣⎧,
𝑣𝑣̅ ≥ 0⎪ 𝐶𝐶𝑣𝑣 �|∆𝑃𝑃𝑖𝑖𝑖𝑖 |(6)
𝑖𝑖 =������
⎨𝑄𝑄
𝑜𝑜𝑜𝑜𝑜𝑜 + ∆𝑄𝑄𝑣𝑣𝑣𝑣𝑣𝑣𝑣𝑣,
𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒⎪⎩ 𝐶𝐶𝑣𝑣 �|∆𝑃𝑃𝑜𝑜𝑜𝑜𝑜𝑜 |4

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where 𝐶𝐶𝑣𝑣 is the flow coefficient of the servovalve, and ∆𝑃𝑃𝑖𝑖𝑖𝑖 , ∆𝑃𝑃𝑜𝑜𝑜𝑜𝑜𝑜 are the pressure drop across the inlet and outlet
edges, respectively of the spool, which can be measured by pressure sensors. The feedback terms ∆𝑄𝑄𝑣𝑣𝑣𝑣𝑣𝑣 , ∆𝑄𝑄𝑣𝑣𝑣𝑣𝑣𝑣𝑣𝑣
are computed by the same equation (5) with different gains. Although not shown in the equations, we may consider
additional flow in advance to compensate leakage rather than relying on feedback only.
Phase 2 (1~2 [s])：The joint J1 needs high positive power, while J2 need high negative power for breaking. SB-mode is suitable forJ1 because single SP cannot provide enough flow. Contrary, N-mode is enough for J2 because the power isprovided by the external load, and the servovalve can easily generate back pressure by throttling.
Therefore SP1 uses Eq.(4), where the feedforward term may hit the maximum flow of the SP. The missing flow
is made up by SP2. The same equation (5) is used, but only the feedback term is used.
Phase 3 (2~3 [s]):The both cylinders stop the motion, but still need high force (anti-gravitational force). Therefore, the robot
should operate at DB-mode as in Phase 1.
Phase 4 (3~4 [s])：The same situation as Phase 2 except for that the polarity of the velocity and load just reverses.
Phase 5 (4~5 [s]):The all SPs stop (in reality, they may rotate at slow speed to compensate for the leakage), and each servovalve
controls the position of each cylinder at (low) common pressure.
Figure 4 shows the operation range of the pressure and flow in this example. Table 1 summarizes the control
law for each phase.

Flow

Q0
NΔQ

2ΔQ
ΔQ
Pressure

0

P0

P0+PbstFigure 4. Operating region of the proposed circuit with two modules during the motion of the weight liftingtask. The curve shows the trajectory of J1 joint only.
Table 1. Operation modes and control

J1

J2

Phase 1, 3

DB-modeDB-modeControl of SPs
SP1 : J1 Position
SP2 : J2 Position

Phase2SB-modeAB-modeSP1 : J1 Position
SP2 : J1 Position

Phase4AB-modeSB-modePhase5

N-modeN-modeSP1 : J2 Position
SP2 : J2 Position
SP1 : Stop
SP2 : Stop

Control of switching valves
V1：-1
V2：-1
V1：0
V2：+1
V1：+1
V2：0
V1：+1
V2：+1

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POSITION TRACKING SIMULATIONThis section shows the simulation result of the position tracking control of the 2-link robot, where the desired task
and trajectory were already illustrated in Fig.3. Note that this task is typical for heavy load manipulators. The
simulation is conducted by combining AMEsim and MATLAB/Simulink. The arm position is controlled bycylinder position feedback through the inverse kinematics of the four-bar linkage. The cylinder has the piston of
the diameter 25 mm, rod of the diameter 12 mm, and 100 mm stroke. The common pressure value is set to 4.0
MPa, which is slightly higher than the pressure for the gravity compensation of the arm itself. For comparison,
we set the following two conditions: (a) using DB-mode only; (b) using all modes. The result is shown in Fig. 5.
Each panel shows, from the top, the position, pressure, and flow of the outlet port of the modules for the J1 (left)
and J2 (right).
The position tracking failed in J1 for the condition (a) during Phase 2. This is because of the shortage of the flow
and hence, pressure. The error is recovered in Phase 3. On the other hand, in condition (b) the inlet flow, hence
the pressure, is timely increased by SB and AB mode during Phase 2, resulting in better tracking performance.
Although the flow of SP2 is transferred to J1 actuator, J2 works well with the throttling control for the negativeload. Therefore, there is no need to supply flow to J2 except for the make-up flow. However, when the phasetransit from Phase 2 to Phase 3 around 2 [s], a small overrunning can be seen in J2 joint due to the transmission
delay of the flow.
We computed the power consumption of this simulation, and compared with a conventional servovalve drive
system of the supply pressure 14 [MPa], that is, the highest pressure required to achieve the same task. The scoreof the proposed system was J1：35.8 [J], J2：21.14 [J], over the conventional one J1：92.88 [J], J2：70.45 [J].
These results show the clear advantage of the proposed circuit not only over the conventional servo systems, butalso against the decoupled load sensing servo systems as we expected.
Joint 1 State
80

Des

Act(a)

60

-80

40
20
0

-90
0

1

2

3

4

15

5

0

1

2

3

4

5

0

1

2

3

4

5

0

1

2

3

4

5

15

(a)
(b)

10

10

5

5
0

1

2

3

0.6

4
(a)

5
0.4

(b)

0.4

0.2

0.2
0

Joint 2 State-70

Act(b)

0

1

2

3Time [s]450Time [s]Figure 5. Simulation results on position tracking of two-link robot arm. The shaded region indicates Phase 2 of the motion
sequence (see Fig.3).

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CONCLUSION
We proposed a new modular hydraulic circuit suitable multi-axis robots. This circuit is characterized by modular,pressure boost, flow summing and high response. We have shown the realization of the circuit and possible
operating modes, then presented the control method for each mode. The proposed circuit was numericallyevaluated on a two-link serial manipulator. The position control result validated the proposed method by its clear
improvement of tracking performance especially in the shared boost mode, where small pumps work together togenerate high pressure and flow. The energy consumption is significantly reduced from the conventional
servovalve system. Notably, the circuit can be installed to the old servovalve systems with little modification
because it is upper compatible. Comparison with existing hydraulic circuits is left for future work. Besides thetheoretical analysis, application to real robotic arm with joint torque control [2] is ongoing.

ACKNOWLEDGMENTS
This work was supported by KAKENHI under Grant 21H01297 (PI: S. Hyon).

REFERENCES
1. HABIBI, S., Design of new high-performance electrohydraulic actuator, IEEE/ASME Trans. Mechatronics, Vol. 5, no. 2,
2000, pp. 158-164.
2. HYON, S., TANIAI, Y., HIRANUMA, K., YASUNAGA, K., MIZUI, H., Overpressure compensation for hydraulic hybrid
servo booster applied to hydraulic manipulator, IEEE Robotics and Automation Letters, Volume: 4, Issue: 2, 2019, pp.942949.
3. JANSON, A., PALMBERG, J., Separate controls of meter-in and meter-out orifices in mobile hydraulic systems, SAE
Transactions, Vol. 99, no. 2, 1990, pp. 377-383.
4. MATTILA, J., TAPIO, V., Energy-efficient motion control of a hydraulic manipulator, IEEE International Conference on
Robotics and Automation, 2000, pp. 3000-3006.
5. YAO, B., LIU, S., Energy-saving control of hydraulic systems with novel programmable valves, 4th World Congress onIntelligent Control and Automation, 2002, pp. 3219-3223.
6. KAI, T., MIZUI, H., HYON, S., Hydraulic Hybrid Servo System H2SB+ for Multi-axis Robots, The 38th Annual
Conference of the Robotics Society of Japan, 2020, 1F2-03.
7. LINJAMA, M., Digital Fluid Power – State of the Art, Proceedings of the 12th Scandinavian International Fluid Power
Conference Vol3, 2011.
8. RUIKAI, B., TAO, W., QINGFENG, W., A multi-pump multi-actuator hydraulic system with on-off valve matrix forenergy saving, IEEE/ASME International Conference on Advanced Intelligent Mechatronics, AIM, 2019, pp. 1110-1115.
9. SHIMIZU, J., OTANI, T., HASHIMOTO, K. and TAKAHASHI, A., Downsizing the motors of biped robot using ahydraulic direct drive system'', IEEE-RAS Humanoids, 2018, pp.580-586.
10. BUSQUETS, E., IVANTYSYNOVA, M., A Multi-actuator displacement-controlled system with pump switching – Astudy of the architecture and actuator-level control, JFPS International, Volume: 8, Issue: 2, 2014, pp. 66-75.
11. HYON, S., SUEWAKA, D. TORII, Y., OKU, N., Design and experimental evaluation of fast torque-controlled hydraulichumanoid robot, IEEE/ASME Transactions on Mechatronics, Vol. 22, no. 2, 2017, pp. 623-634.

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Comparison of Mechanical Drive System andHydraulic Direct-Drive System for Motor Power
Juri SHIMIZU*, Takuya OTANI**, Kenji HASHIMOTO***, Atsuo TAKANISHI*** The Graduate School of Advanced Science and Engineering,
Waseda University, and Hitachi, Ltd.
832-2 Horiguchi, Hitachinaka , Ibaraki, 312-0034 Japan(E-mail: juri-shimizu1112@asagi.waseda.jp)** The Department of Modern Mechanical Engineering, Waseda Un iversity
17 Kikui-cho, Shinjuku-ku, Tokyo, 162-0044 Japan*** The School of Science and Technology, Meiji University
1-1-1 Higashimita, Tama -ku, Kawasaki, Kanagawa, 214-8571 Japan

Abstract. In biped robots with serial links driven by an electric motor, a motor and transmission are installed in eachjoint, causing the legs to become very heavy. Therefore, a high-power motor is required at the root of each leg. To
overcome this problem, the use of a hydraulic system has been proposed since it would reduce the required output at
the root joint by placing the relatively heavy pressure source at the root and adopting lightweight cylinders at eachjoint. We previously proposed a hydraulic direct-drive system (HDD) based on a flow-based control system by apump that uses a single-rod cylinder for biped humanoid robots. In this study, an HDD with an interlocking circuit isapplied to a walking simulation. The simulation results demonstrate that the motor output of the proposed system iscomparable to that of the mechanical drive system.
Keywords: Hydraulics, Humanoid, Biped locomotion, Flow-based control, Simulation

INTRODUCTION
The products people use need to undergo safe quantitative assessments. However, the main method currently
used for this purpose (user rating) provides only qualitative results. Therefore, our research group proposes
using biped humanoid robots instead of humans to evaluate product s. WABIAN-2R (WAseda Biped
humANoid-No. 2 Refined) developed by our group can walk with its knees extended by introducing a pelvic
mechanism as shown in Fig. 1 [1]. In addition to WABIAN-2R, we have developed a robot that can evaluatevarious products. It runs as well as walks. The robot requires a high-power actuator in order to move, but due tothe robot’s size limitation, a high-power electric motor is d ifficult to implement. To solve this p roblem, we
proposed a method to generate a large torque by using the vibration of the pelvis and the elasticity of the legs [2,3].
Urata et al. developed a method to improve the continuous output torque in a liquid coolin g system in order toincrease the output of an electric motor [4]. Other robots have two or three motors on the drive shaft [5 –7].
These approaches have led to the development of high speed, high torque, and highly mobile joints forhumanoid legs.

Figure 1. WABIAN-2R

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The hydraulic system can realize a robot that is the sa me size as a human. For example, Boston Dynamics
developed ATLAS [8]. Hyon et al. developed a hydraulic humanoid robot [9] called T aeMu, which has the sameleg link ratio and weight dist ribution as humans. TaeMu can directly control the torque of each joint by
adjusting the pressure of the actuator by using a servo valve. However, in its intended use, TaeMu needs toreproduce a predetermined motion pattern in order to evaluate a product. This requires a system that can directly
control the joint angle.
Previously, we proposed a hydraulic direct-drive system (HDD) based on a displacement control system thatuses a single -rod cylinder for a biped humanoid robot [10, 11]. In HDD, the meter-in flow rate of a single-rod
cylinder is controlled by the pump discharge flow rate, and the cylinder meter-out flow rate is controlled by aproportional valve. The proposed system showed excellent energy transfer efficiency and excellent speed
tracking capability in the evaluation bench system. In addition, the results show t hat bipedism can be achieved
with an HDD that uses a collaborative simulation model of hydraulics and machines [12].
However, the HDD is mounted individually at each joint. This caused the problem of needing to increase theoutput of each pump to meet the maximum required output of each joint. To so lve this problem, an interlocking
circuit was proposed that focuses on the symmetry structure of the vibe mechanism and the symmetry of motion[13]. The proposed system reduced the maximum motor output by 27.3% in an evaluation bench device thatsimulated the load during walking compared with the case without the system.
In addition, comparing the performance in a walking simulation with the performance of a stand-alone HDD, wefind that the interlocked HDD reduces the output power of the motor by 34.1% [14].
In this study, HDD with an interlocking circu it and conventional mechanical drive system is applied to a bipedhumanoid robot simulation model. The results of the performan ce comparison with conventional mechanicaltransmission verify the effectiveness of the proposed interlocked HDD.
The remainder of this paper is organized as follows: Sect ion 2 introduces the features of the conventional
mechanical drive system and the interlocked HDD for biped humanoid robots. Section 3 describes theexperimental setup, procedures, and results. Section 4 discusses future prospects and concludes the paper.

COMPARISON OF DRIVE SYSTEMSThis sect ion describes the features of the mechanical drive system that has been implemented in our bipedhumanoid robots and the proposed HDD.

Mechanical drive system
As shown in Fig. 1, our developed WABIAN-2R amplifies the motor torque by a harmonic drive connected by atiming belt to achieve high torque [1]. The harmonic drive has a high reduction gear ratio and no backlash. For
example, the WABIAN-2R h ip joint pitch axis exerted torque with a harmonic drive with a reduction ratio of
1/100. Although high torque and precise control can be achieved, the harmonic drive tends to have low backdrivability, which is the ease of following an operation against an external force from the load side, andtransmission efficiency is also lower than that of general gear systems.
Fig. 2 shows the characteristics of the transmission efficiency of the harmonic drive SHD17 used in the hip p itch
joint of WABIAN-2R [15]. The harmonic drive ha s the characteristic that efficiency deteriorates at the lower
rated torque ratio because there is mechanical loss due to dragging torque caused by friction.

Figure 2. Efficiency of the harmonic drive (SHD17)

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Hydraulic direct-drive system
This sect ion describes the features of the hydraulic system and the mechanical link for our proposed HDD for
biped humanoid robots.
Hydraulic system
Our proposed HDD controls the flow rate of the actuator by pump derived flow. The proposed circuit is shownin Fig.3. The drive velocity and direction of the cylinder can be controlled by the pump d erived flow rate anddirection. The cylinder meter outflow rate can be controlled by the proportional valves installed between thepump and the cylinder in a negative load situation when the cylinder drive d irection and load direction are thesame.
Furthermore, by providing a valve that connects two pump outputs and two valves that connect two cylinders, it
is possible to concentrate the pump derived flow rate on the cylinder with a higher load and drive the cylinder
with a lower load by the discharge flow rate of the higher load cylinder. The interlocking circu it can reduce theoutput power of the motor for the pumps.
In this study, Valve 3c that connects the cap side of Cylinders 1 and 2 is added to a previously proposed circuit[13]. A method has already been proposed to control Valve 3b, which connects the rod-side cylinder [13]. Valve3c can also be controlled on the basis of the proposed control method by reversing the relationship between themeter-in flow and meter-out flow of Cylinders 1 and 2 at an interlocking mode.
The conversion efficiency from the motor output to the cylinder output of the proposed system is determined by
the efficiency of the pump shown in Fig. 4. As shown in Fig. 4, high pump revolution and high pump dischargepressure improve pump efficiency. Therefore, in this study, we aim to improve the efficiency of the hydraulicsystem by increasing the pump revolution by reducing the pump displacement and increasing the pressure by
reducing the cylinder diameter.

Figure 3. Proposed hydraulic circuit with interlocking drive circuit

Figure 4. Efficiency of the pump (TFH-080-U-SV)

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Mechanical link
The proposed system uses a cylinder as a highly efficient hydraulic actuator. The cylinder is linearly driven, andto use it to drive the rotary joints of biped humanoid robots, a link mechanism is used to convert linear motion
into rotary motion. The proposed system can directly control the cylinder speed by the pump revolution. The
cylinder speed is easier to control if the joint angular velocity can be determined by the pump revolution. To
achieve this, the linear stroke and rotation angle must b e linear. Therefore, a linear characteristic was obtainedby optimizing the parameters using the 4-bar linkage mechanism shown in Fig. 5 [11]. Table 1 shows theactually designed link parameters. Fig. 6 (a) shows the relationship between the cylinder stroke and the joint
angle.
A simulation showed that walking can be achieved by the feedforward flow rate control by pump andproportional valves and the simple proportional control that compensates for the error with the target angle bythis linkage mechanism [12].
To improve the efficiency of the drive system, the conversion efficiency of the cylinder output force and joint
torque should be focused on. Fig. 6 (b) shows the relationship between cylinder stroke and joint torque at acylinder output force of 100N. This characteristic indicates that the conversion rate from the cylinder output
force to joint torque varies greatly depending on the cylinder stroke.

Figure 5. Four-bar linkage mechanisms
Table 1. Four-bar linkage mechanisms parameters
119
Link 1 [mm]113
Link 2 [mm]59
X A [mm]
34
ZA [mm]100
X B [mm]
300
ZB [mm]0
X C [mm]
57
ZC [mm](a) Cylinder stroke and joint angle(b) Cylinder stroke and joint torque (Cylinder output: 100 [N])Figure 6. Link characteristics

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Evaluation
The required motor powers of the proposed interlocked HDD and conventional mechanical drive sy stem werecompared experimentally. The mechanical hydraulic co-simulation model was used to determine whether the
maximum output of the motors can be reduced by the walking pattern. The maximum output of the motor during
walking with the mechanical drive system was compared with the situation when the HDD and interlocked
HDD were applied. The details are presented in the next section.

SIMULATION
Simulation setup
To compare the proposed HDD with the conventional machine drive system, we developed a simulation modelof a biped humanoid robot that implements each system. We used the physical modeling tool LMS Imagine.Lab
Amesim™ (Siemens K.K.) to evaluate the proposed system.
Mechanical Drive system
A model for evaluating the motor output of a mechanical drive system was constructed using the threedimensional mechanical analysis lib rary. The construction of the mechanical model was based on the linkmechanism and mass distribution of WABIAN-2R [1]. Since the proposed hydraulic system is applied to the hip
pitch joint, we applied the harmonic drive efficiency shown in Fig. 1 to this joint. In the constructed model, theefficiency was determined from the motor revolutions and torque, and the motor output power was calculated byconsidering this efficiency with respect to the joint output.
Hydraulic direct-drive system
To evaluate the proposed interlocked HDD, we have developed a mechanical hydraulic co-simulation model ofa biped humanoid robot. As the mechanical model, the model constructed for the mechanical drive system wasused. This mechanism model and the proposed interlock ed HDD hydraulic model were coupled. Our proposed
hydraulic system was applied to the hip pitch joint. The hip joint required high torque and joint density, andspace limitations make it difficult to install large drive systems. The other joints in the model were the same as
those in the mechanical drive system. Fig. 7 shows a co-simulation model of hydraulics and machinery, and Fig.
8 shows the details of a hydraulic model. The hydraulic model was based on the circuit shown in Fig. 3 andreflects the valve characteristics and motor responsiveness based on the experiments performed in a prev ious
study [11]. Furthermore, for the efficiency of the pump, the characteristics shown in Fig. 4 were implemented.
The corresponding parameters are shown in Table 2. The four-bar linkage mechanism installation angle wasadjusted so that the maximum torque shown in Fig. 6 (b) could be exerted in the joint range used in the walking
pattern performed in the experiment.
Table 2. Hydraulic circuit model parameters

Pump displacement [cc/rev]
Pump relief pressure [MPa]Cylinder stroke [mm]Cylinder piston diameter [mm]Cylinder rod diameter [mm]Variable relief valve set pressure [MPa]Hose inner diameter [mm]Pump to Valves 1a, 1b, 2a, 2b
Hose length Pump to Valve 3a[mm]Valves 1a, 1b, 2a, 2b, to Cylinder 1, 2
Valves 3b and 3c to Cylinder 1, 2
Valves 1a, 1b, 2a, 2b : Nominal flow rate [L/min/MPa]
Valves 3a, 3b, 3c : Nominal flow rate [L/min/MPa]
40 [deg C]
Oil kinematic viscosity [mm 2 /s]
100 [deg C]
Experimental oil temperature [deg C]

0.8
21
132
15
10
15
6.4
1000
300
2000
300
5
10
14.7
3.7
25

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Figure 7. Co-simulation model

Figure 8. Hydraulic model

Simulation procedure
First, a walking simulation was executed when the mechanical drive system and HDD were applied. From theresults, the phase in which the motor output increased during walking was determined. If the output of the motorfor driving one leg is h igh and the output of the motor on the opposite side is low, the output of the motor shouldbe able to be suppressed by using an interlocking circuit.
From the analysis results, the timin g of applying the interlocking circuit was determined, and a walking
simulation was performed. In addition, the results of the gait simulation will be used to evaluate how much theproposed interlocked HDD reduces the motor output. To verify the effectiveness of the proposed system, we
compared it with the results of the walkin g simulation of the mechanical drive system reported in a prev ious
study [1].
By reducing the output of the motor, the motor can be shrunk. The motor is selected on the basis of the required
output. The goal is to achieve the maximum output required for each axis. If the peak output of each motor is
reduced by the interlocking method, a smaller motor can be selected, which will contribute to the reducing theweight of the robot. This means that by installing a motor with the same output as before, higher performance
than before can be achieved. Therefore, the comparative experiment evaluated the peak output of the motor
during walking.
The walking pattern of the simulation experiment was generated using the offline pattern generation method
proposed by our research group. The group has a proven track record of walkin g with WABIAN-2R [1], the6

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basis of the mechanism model. In this study, we applied the walking pattern with knee bending, which is awalking pattern of most biped robots.

Simulation results
Fig. 9 shows the simulation results for the walkin g motion of the biped humanoid robot model with the HDD.
The results show that the proposed HDD can follow the specified trajectory under the walking motion load,
thereby achieving biped walking.
Fig. 10 shows the power of Motors 1 and 2 with the HDD and mechanical drive system. As shown in Fig. 10,
the power of Motor 1 for the right leg increased at a single stance left phase, such as at 2.7–3.4. In this phase,
Motor 1 must raise the right leg quick ly. Thus, the output of M otor 1 increased, as shown in Fig. 10. The
maximum power was 177.8 W in the HDD and 87.7 W in the mechanical drive system.

Figure 9. Walking motion of co-simulation model(a) Right leg side(b) Left leg side
Figure 10. Motor output power

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The 11th JFPS International Symposium on Fluid Power 2020(a) Right leg side(b) Left leg side
Figure 11. Motor output power(a) Flow rate(b) Pressure
Figure 12. Pump1 port B condition

Fig. 11 shows the power of Motors 1 and 2 when the interlocking drive mode was applied. This mode wasapplied during the sin gle stance phase, and outputs of Pumps 1 and 2 connect to the swing leg side cylinder by

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opening Valves 3a and 3c. As shown in Fig. 11, the peak power of Motor 1 was reduced to 89.7 W. Therefore,
the maximum motor power can be comparable to that of the mechanical drive system.
Fig. 12 shows Pump1 port B the flowrate and the pressure with the HDD and the interlocked HDD. As shown in
Fig. 12, although the pressure was comparable, the flowrate decreased by interloking system at a single stanceleft phase, such as at 2.7–3.4. Therefore, the peak power of Motor 1 at the phase shown in Fig. 11 (a) was
reduced compared to the result shown in Fig 10 (a).

Discussion
As shown in Fig. 10, when comparing the mechanical drive system and a single HDD, the HDD had a higher
motor output. This is not a problem with the hydraulic system, but a problem with the link mechanism. Fig . 13
compares the output of the joint when the efficiency of the harmonic drive is not taken into consideration andthe output of the cylinder when the efficiency of the hydraulic circuit is not taken into consideration. The
maximum outputs of the joint and cylinder were 58.4 and 125.8 W, respectively.
As shown in Fig. 13, the cylinder that drives the joint via the link mechanism required a larger output even when
the same walk was performed. This can be improved by searching for energy -efficient link parameters by
optimizing the link mechanism. However, since there is a trade-off between the cylinder stroke and the linear
characteristics of the joint angle shown in Fig. 6 (a), the optimum link ratio needs to be determined from theviewpoints of controllability and efficiency. In addition, if the hydraulic motor has higher efficiency than thelink, an approach should be considered that converts linear motion into rotary motion by a method other thanlinking.
However, even in that case, the effect of reducing the maximum motor output by the interlocking circu it does
not change. Therefore, the effectiveness of the proposed system is not lost. In addition, it would be effective tohave an approach that maximizes the effect of the interlockin g circu it by controllin g the output peaks of the left
and right legs to shift by creating a walking pattern.

(a) Right leg side(b) Left leg side
Figure 13. Actuator output power

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CONCLUSION AND FUTURE WORKSIn this study, the performances of a hydraulic interlocking drive system with a conventional hydraulic d irectdrive system (HDD) and a mechanical drive system were compared to verify the motor output power. A
hydraulic and mechanical co-simulation model was used for the evaluation. Simulation results showed that the
motor output peaked during the swin g leg in walking, and the timing of interlocking was adjusted on the basis of
the experimental results.
The results also showed that the motor output in the interlocked HDD could be reduced to the level of amechanical drive system and that the proposed system contributes to reducing the size of the installed motors forbiped humanoid robots.
In the future, linkage mechanism efficiency will need to be improved for energy saving, and the proposed HDD
will be applied to a biped humanoid robot to conduct a walking experiment.

ACKNOWLEDGMENTS
This study was conducted with the support of the Research Institute for Science and Engineering of Waseda
University and the Future Robotics Organization of Waseda University. Th is study was part of the humanoid
project at the Humanoid Robotics Institute of Waseda University.

REFERENCES
1. OGURA, Y., SHIMOMURA, K., KONDO, H., MORISHIMA, A., OKUBO, T., MOMOKI, S., LIM, H., TAKANISHI,
A., Human-like Walking with Knee Stretched, Heel-Contact and Toe-Off Motion by a Humanoid Robot, IEEE/RSJ Int
Conf on Intelligent Robots and Systems, 2006, p.3976–3981
2. OTANI, T., HASHIMOTO, K., HAMAMOTO, S., MIYAMAE, S., SAKAGUCHI, M., KAWAKAMI, Y., LIM, H.,
TAKANISHI, A., Knee Joint Mechanis m that Mimics Elastic Characteristics and Bending in Human Running,
IEEE/RSJ Int Conf on Intelligent Robots and Systems, 2015, p.5156–5161
3. OTANI, T., HASHIMOTO, K., MIYAMAE, S., UETA, H., SAKAGUCHI, M., KAWAKAMI, Y., LIM, H.,
TAKANISHI, A., Angular Momentum Compensation in Yaw Direction Using Upper Body Based on Human Running,
IEEE Int Conf on Robotics and Automation, 2017, p.4768–4775
4. URATA, J., NAKANISHI, Y., OKUDA, K., INABA, M., Design of High Torque and High Speed Leg Module for HighPower Humanoid, IEEE/RSJ Int Conf on Intelligent Robots and Systems, 2010, p.4497–4502.
5. ITO, Y., NOZAWA, S., URATA, J., NAKAOKA, T., KOBAYASHI, K., NAKANISHI, Y., OKADA, K., INABA, M.,
Development and Verification of Life-Size Humanoid with High-Output Actuation System, IEEE Int Conf on Robotics
and Automation, 2014, p.3433–3438.
6. PARK, I., KIM, J., LEE, J., OH, J., Mechanical Design of the Humanoid Robot Platform, HUBO, Advanced Robotics,
Vol.21, No.11, 2007, p.1305–1322.
7. KANEKO, K., KAMINAGA, H., SAKAGUCHI, T., KAJITA, S., MORISAWA, M., KUMAGAI, I., KANEHIRO, F.,
Humanoid Robot HRP-5P: An Electrically Actuated Humanoid Robot with High Power and Wide Range Joints , IEEE
Robotics and Automation Letters, Vol.4, No.2, 2019, p.1431–1438.
8. https://www.bostondynamics.com/atlas, (accessed on 1, June, 2021)
9. HYON, S., SUEWAKA, D., TORII, Y., OKU, N., Design and Experimental Evaluation of a Fast Torque-Controlled
Hydraulic Humanoid Robot. IEEE/ASME Transactions on Mechatronics, Vol.22, No.2, 2017, p.623–634
10. SHIMIZU, J., OTANI, T., MIZUKAMI, H., HASHIMOTO, K., TAKANISHI, A., Experimental Validation of HighEfficiency Hydraulic Direct-Drive System for a Biped Humanoid Robot, IEEE Int Conf. on Robotics and Automation,
2019, p.9453–9458.
11. SHIMIZU, J., OTANI, T., HASHIMOTO, K., TAKANISHI, A., Evaluation of Tracking Control for Hydraulic Directdrive System, JFPS International Journal of Fluid Power System, Vol.13, No. 3, 2020, p.17-24.
12. SHIMIZU, J., OTANI, T., HASHIMOTO, K., TAKANISHI, A., Simulation of a Hydraulic Direct-Drive System for a
Biped Walking Robot, 16th Scandinavian Int Conf on Fluid Power, 2019, B6.1.
13. SHIMIZU, J., OTANI, T., MIZUKAMI, H., HASHIMOTO, K., TAKANISHI, A., Experimental Validation of HydraulicInterlocking Drive System for Biped Humanoid Robot, IEEE/RSJ Int Conf on Intelligent Robots and Systems, 2019,
p.7163–7169.
14. SHIMIZU, J., OTANI, T., HASHIMOTO, K., TAKANISHI, A., Simulation of an Interlocking Hydraulic Direct-Drive
System for a Biped Walking Robot, 12th International Fluid Power Conference, 2020, p.211-220.
15. http://www.takako-inc.com (accessed on 1 June, 20219)
16. https://hds-tech.jp/pdf/hd01_sougou_0204.pdf (accessed on 1 June, 2021)

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Genaral Session | Oil hydraulics

Energy Saving[GS2-01] Research on the Characteristics of the Cylinder Exhaust-Return EnergySaving System〇Yuto Fujiwara1, Mitsuru Senoo1, Hiroyuki Asahara1 （1.SMC Corporation）[GS2-02] Design Guideline and Investigation of Accumulator Parameters for a Novel
Hybrid Architecture〇Seiji Hijikata1, Kazuhisa Ito2, Hubertus Murrenhoff1 （1.Institute for Fluid Power Drives andSystems (ifas), RWTH Aachen University, 2.Department of Machinery and Control Systems,
College of Systems Engineering and Science Shibaura Institute of Technology）[GS2-03] Research Regarding the Energy Saving Conditions of the Air Blow for
Removing and Drying Out Water〇Daisuke Kurakami1, Keiichirou Koga1, Gohei Harimoto1 （1.SMC Corporation）[GS2-04] Energy Regeneration and Reuse of Excavator Boom System with HydraulicConstantly Variable Powertrain〇Cuong Tri Do1, Kyoung Kwan Ahn1 （1.University of Ulsan）[GS2-05] Efficient Closed Pump Controlled Hydraulic-Gas Balanced Energy
Recovery Driving Method for Hydraulic Excavator Boom
〇Lianpeng Xia1, Long Quan1, Hongjuan Zhang1, Yunxiao Hao1, Lei Ge1 （1.Taiyuan University ofTechnology）[GS2-06] Experimental Validation of Improvement of the Overall Efficiency forElectro-hydraulic Drive System using Efficiency Maps
〇Ha Tham Phan1, Yasukazu Sato1 （1.Yokohama National University）©The Japan Fluid Power System SocietyGS2-01

The 11th JFPS International Symposium on Fluid Power 2020

Research on the Characteristics of the Cylinder Exhaust-Return
Energy-Saving System
Yuto FUJIWARA*, Mitsuru SENOO*, Hiroyuki ASAHARA** SMC Corporation Tsukuba Technical Center, 4-2-2,
Kinunodai, Tsukubamirai-shi, Ibaraki-ken, 300-2493 Japan(E-mail: fujiwara.yuto@smcjpn.co.jp)

Abstract. In recent years, as part of efforts toward SDGs (Sustainable Development Goals), energy-saving orientation
has become stronger. In addition, it is said that compressed air generation is responsible for approximately 20% of thepower consumption in a factory, so energy-saving systems are required to reduce the air consumption of pneumatic
cylinders used in various applications. Therefore, the authors have developed an exhaust-return energy-saving system
as a new energy-saving actuation system. This system reuses the compressed air normally exhausted after the end of
the working stroke, as an air pressure source during the return stroke of the cylinder drive. In other words, nocompressed air is consumed in return stroke, so approximately 50% of the air consumption can be reduced per
operating cycle. This paper introduces this system and describe the results of calculating the characteristics of thesystem by simulation.
Keywords: Energy-Saving, Pneumatic System, Exhaust-Return System, Simulation, Characteristic Calculation.

INTRODUCTION
In recent years, as part of efforts toward SDGs (Sustainable Development Goals), energy-saving orientation has
become stronger. In addition, it is said that compressed air generation is responsible for approximately 20% of thepower consumption in a factory, so energy-saving systems are required to reduce the air consumption of pneumatic
cylinders used in various applications.
Normally, in cylinder drive, the same supply pressure is supplied and exhausted in an operating cycle. However,in general applications such as transport and clamping, only the working stroke requires thrust, and a low pressure
is often sufficient for the return stroke. Therefore, in order not to impede practicality, the mainstream of energysaving was the return stroke that does not require thrust. For example, two-pressure actuation and double power
operation mechanisms were adopted in accordance with the energy-saving actuation system [1] for reducing thethrust of the cylinder.
However, when pursuing energy-saving by reducing the thrust, there is a limit to the reduction because it isnecessary to supply at least enough compressed air to the cylinder to satisfy the required thrust.
Therefore, the authors have developed an exhaust-return energy-saving system as a new energy-saving actuation
system. This system reuses the compressed air normally exhausted after the end of the working stroke, as an airpressure source during the return stroke of the cylinder drive. In other words, no compressed air is consumed inreturn stroke, so approximately 50% of the air consumption can be reduced per operating cycle.
This paper introduces exhaust-return energy-saving system, and describe the results of calculating thecharacteristics of the system by simulation.

EXHAUST-RETURN ENERGY-SAVING SYSTEM
This system reuses the compressed air supplied to the working stroke that should be exhausted to the return stroke
and reuses it for the return stroke as an air pressure source. Therefore, the return stroke does not require an
additional compressed air supply. In other words, no compressed air is consumed in return stroke, so
approximately 50% of the air consumption can be reduced per operating cycle. The thrust of the working stroke
does not change.
As shown in Figure 1, the head side and rod side are connected through the check valve and solenoid valve. The
flow path connecting the head side flow path and the rod side flow path is called a return flow path.
In the working stroke, compressed air is supplied to the head side and exhausted from the rod side the same as thegeneral circuit.

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Figure 1. Compressed Air Flow During Operation.

On the other hand, on the exhaust-return stroke, the compressed air supplied to the head side flows as shown
below so it can be reused.
(1) Return flow path: The compressed air supplied to the head side is returned to the rod side through the checkvalve.
(2) Rod side flow path: The returned air is reused and redirected to supply pressure to the rod side port.
(3) Head side flow path: At the same time as the return, the compressed air supplied to the head side is exhausted
to the atmosphere and the head side pressure drops.
With this operation, the piston operates with the pressure difference between the rod side and head side. Thus theamount of compressed air supplied in the return stroke is zero.

MODELING and VALIDATION
In order to clarify the characteristics of this system, a simulation model was created. Figure 2 shows thesimulation model for the exhaust-return stroke, and Table 1 shows the parameters used for the calculation. For
the sake of calculation simplicity, state change of air is assumed to be isothermal change.

Figure 2. Simulation Model (Exhaust-Return Stroke)2

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Parameter
P
V
C
b
A
θ
Ff
mp
mL
μ
T
v

t
R
ρ

g

Table 1. Simulation Parameters
Unit
Absolute pressure
MPa
Volumem3
Sonic conductancedm3 / ( s･bar)
Critical pressure ratio
Piston aream2
Mounting angle
radFriction force of piston rod
N
Mass of cylinder movable parts
kg
Load masskg
Friction coefficient
Temperature
K
Velocitym/sTimes
Gas constant ( = 287)
J / (kg･K)
Standard reference atmosphere density ( = 1.185) kg / cm3
Gravitational acceleration ( = 9.8)m / s2

Name

1
2
3
4
5
h
r
s
k

v

Subscript
Cylinder head side
Cylinder rod side
Head side flow path
Rod side flow path
Return flow path
Head side port
Rod side port
Speed controller
Check valve
Solenoid valve

Based on this model, the flow of compressed air between each volume is calculated.
For example, the mass flow rate G13 flowing from the cylinder head side to the head side flow path isrepresented by the flow rate equation (1), and the puressure change dP1 in the cylinder head side pressure due tothis is represented by the state equation (2). Then, from the obtained pressure difference in the cylinder, thevelocity change dv of the cylinder is represented by the equation of motion (3).

G 13 =G 13 =

ρ C h P1
100

ρ C h P1
100

1−

( PP13 − b h ) 2(1 − b h )

2

293
.
T

[ PP13 > b h : subsonic flow]
(1)

293
.
T

[ PP13 ≤ b h : choked flow ]

dP1
= −
dt

( RTG 13 + P1
V1

dV 1
)
dt .

P2 A 2 − P1 A1 − F f + ( m p + m L ) g sin θ − µ m L g cos θ
dv
=
.
dt
m p + mL

(2)

(3)

These parameters are repeatedly calculated at time step to reproduce the behavior of the system.
Figure 3 shows the results of confirming the reproducibility of the model for the experiment, and Table 2 shows
the conditions for the experimental and calculation.

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Figure 3. Reproduction of the Simulation Model (Exhaust-Return Stroke).

Table 2. Experimental and Calculation Conditions
ConditionNameØ 20 mm – 25 mm
Cylinder ( bore – stroke )Ø 4 mm – 2000 mm
Head side flow path ( I.D. – length )Ø 4 mm – 2000 mm
Rod side flow path (I.D. – length )Ø 4 mm – 100 mm
Return flow path (I.D. – length)
0.7 MPa G
Supply pressure
0 rad ( horizontal )Mounting angle
11 kg
Load mass
Since the simulation results were in good consistency with the experimental results, the validity of the model
was confirmed. Therefore, the characteristics of the system will be clarified using this model.

SIMULATION and DISCUSSION
This system reuses compressed air already supplied into the system compressed air instead of supplying newcompressed air, if the return air pressure is not sufficient, operation may not acceptable for the application.
Therefore, it is necessary to estimate the remaining pressure after the stroke (hereinafter called return pressure),
and the optimum conditions were calculated by simulation.
As a result, it was clarified that two factors affect the return pressure.
The first factor is control of the exhaust. The exhaust-return stroke velocity should be adjusted with a meter out
speed controller. Therefore, the larger the exhaust flow rate by increasing the opening in the speed controller, thefaster the exhaust-return stroke velocity will be, however, the amount of compressed air flowing from the branch
to the return flow path is reduced, which decreases the return pressure. This means that in the exhaust-return stroke,
the velocity and thrust are in an inverse correlation relationship as shown in Figure 4.

Figure 4. Relationship between the Exhaust-return stroke velocity and return pressure.

The second factor is the volume ratio before and after the exhaust-return stroke. Here, it is the ratio of "the volume
that stores the pressure supplied after the working stroke (V1 + V3 + V5) " to "the volume in which the returned airflows in during the exhaust-return stroke (V2 + V4 + V5) ".
If this ratio is too small, the amount of compressed air that can be returned will be reduced and sufficient pressure
will not be maintained during the stroke. On the other hand, if the ratio is too large, the returned air expands andthe pressure drops.
From these two factors, the characteristics of the return pressure are shown in Figure 5.

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Figure 5. Characteristics of the Return Pressure
As shown in Figure 5, the return pressure peaks at a volume ratio of approximately 0.5. Then, it can be seen that
it is necessary to be within a certain volume ratio range in order to satisfy the required pressure.
For example, if the required pressure is 0.1 MPa G, the availability of exhaust-return stroke is as follows:- If the volume ratio is 3, it is available at all openings.
- If the volume ratio is 5.2, the speed control is limited because it is not available with openings of 6 or more.
- If the volume ratio is 6 or more, it is not available at all openings.
From this simulation result, it was clarified that the characteristics of the exhaust-return energy-saving system areaffected by the volume ratio and speed control, and that they should be used under appropriate conditions.

CONCLUSION(1) The exhaust-return energy-saving system can save approximately 50% of the air consumption byrecycling the exhausted air.
(2) Calculation was performed with the simulation model, which turned out that the volume should be in an
appropriate range and the relationship between the exhaust-return stroke velocity and return pressure areinverse correlation relationship.
The product development and selection software will be established by taking advantage of these factors.

REFERENCES
1. ONEYAMA, N., Energy Saving for Pneumatic System, 2003, pp.289-309.

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Design Guideline and Investigation of Accumulator Parameters for a
Novel Hybrid Architecture
Seiji HIJIKATA*, Kazuhisa ITO** and Hubertus MURRENHOFF**

Institute for Fluid Power Drives and Systems (ifas), RWTH Aachen University
Campus-Boulevard 30, 52074 Aachen, Germany(E-mail: Seiji.Hijikata@ifas.rwth-aachen.de)**
Department of Machinery and Control Systems, College of Systems Engineering and Science
Shibaura Institute of Technology, 307 Fukasaku, Minuma-ku, Saitama-shi, 337-8570 Japan

Abstract. In order to enhance energy efficiency of excavators for global environment conservation, an Open Center
System (OC-System) as a power source has been improved in the world. However, total system efficiency including
the internal combustion engine (ICE) has not been thoroughly considered. On the other hand, a Constant Pressure
System (CP-System) enabling the engine to be driven optimally is developed but is not accepted in the industry due tothe complexity of the required components. Thus, in this research, a hybrid system combining an OC-System with a
CP-System is proposed by using a proposed design guideline and analyzing measurement data of a dig and dump cycle.
It is indispensable to consider the nominal gas volume and pressure level for an accumulator in terms of energy saving
and initial cost. Therefore, the influences of accumulator volume and pressure level are also discussed in this paper.
Keywords: Hydraulic hybrid, accumulator, mobile hydraulics, excavators

INTRODUCTION
To improve the efficiency of excavators many types of different hydraulic architectures have been developed
during recent years [1-5]. One of the most common valve controlled architectures is an Open Center System (OCSystem). In conventional OC-Systems, however, throttling losses are unavoidable when a single pump supplies
fluid to some actuators simultaneously due to the mismatch of the required pressure levels. To eliminate theselosses, a Three Pump Open Center System (TPOC-System) based on an OC-System was developed [6]. Thissystem is shown in figure 1. Using three pumps all three hydraulic actuators are powered by individual pumps,
and therefore throttling losses are almost negligible if flow rate from the pumps to the tank is small enough byfully tilted joystick operation. Another advantage of this system is simple configuration with hydraulic valvescommonly used for excavators. However, in this system, the internal combustion engine (ICE) is not driven
optimally resulting in decreased total energy efficiency.

Figure 1. Outline of hydraulic system for TPOC-System

To improve the total efficiency ifas at RWTH Aachen University proposed a Constant Pressure System (CPSystem), called STEAM [7, 8], see figure 2. This system consists of a large number of switching valves, two

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accumulators used for driving actuators and a pump for charging the accumulators. The most important feature isthat the rotational speed of the ICE is fixed in its high-efficiency region. Moreover, by using switching valves,installed at the piston and rod side of a cylinder and connected to the accumulators, this system can generate powerby utilizing different cylinder forces which contribute to the reduction of throttling losses and recuperate actuator
energy. There are, however, some disadvantages; since a high number of switching valves are needed, widespread
application in the industry is prevented.

Figure 2. STEAM-System

To solve this problem, the authors propose a new hybrid system combining an OC-System and a CP-System [911]. However, previous papers were only discussed for a levelling cycle. Therefore, including a dig and dump
cycle which is the most standard cycle for excavators, this paper proposes a design guideline to develop the new
hybrid system. In addition, the influence of the nominal gas volume and pressure level for an accumulator is
discussed because it is expected that accumulator parameters affect energy saving and initial cost for the new
hybrid system. In this research the TPOC-System is used as a reference system to compare fuel consumption.

NEW HYBRID ARCHITECTURE
This system is designed based on three basic principles from the perspective of energy efficiency and a simple
configuration. First, to design the simple hydraulic circuit of the new system, open center valves are used as abasic hydraulic system, see figure 3. The new system consists of open center valves, two accumulators, which are
set to medium pressure (MP) and high pressure (HP), and a minimum of required components which are used as
constant pressure valves. Second, boom potential energy and swing brake energy can be recuperated using MP
and HP, and they can be provided to actuators. Thus, the power of the pumps for this system can be reduced byusing this stored energy.

Figure 3. Outline of new hybrid system

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Third, the ICE can be operated in the high efficiency region like the STEAM-System. For explanation, a simple
relative efficiency map of the ICE is represented in figure 4. The horizontal axis shows ICE’s rotational speed 𝑛,
and the vertical axis indicates ICE’s torque 𝑇. The values in figure 4 means relative efficiency which is the ratioof total to ICE’s maximum efficiency. Generally, the high efficiency region extensively appears at lower rotational
speed than that are used in today’s conventional excavators. Moreover, the ICE’s friction depending on rotational
speed can be reduced at lower rotational speed than higher speed. Therefore, the ICE is set to a low rotational
speed. The reduction of the ICE’s power resulting from altering the high rotation speed into the low rotation speed
is compensated by the accumulators. Next, the design guideline to develop the minimum of required components
is discussed.

Figure 4. Schematic relative efficiency map of ICE

DESIGN GUIDELINE OF NEW SYSTEM
In order to explain the design guideline of the new hybrid system, the excavator’s actuator status is explained at
first. Figure 5 describes the load situation for the actuators. The x-axis shows the flow required by each actuator(𝑄L ), the y-axis shows load pressure (𝑝L ) of the actuator. Herein 𝐹L is the load force, and 𝑣 is the actuator speed.
𝑝B is the actuator pressure of the cylinder’s bottom side, and 𝑝R is the actuator pressure of the cylinder’s rod side.
𝐴B is the area of the cylinder’s bottom side, and 𝐴R is the area of the cylinder’s rod side. 𝛼 is the actuator area
ratio which is 𝐴R / 𝐴B for the cylinder and 1 for the swing. Depending on the movement, each actuator experiences
either a resistive force opposing its motion (Q I and Q III) or an assistive force aiding its motion (Q II and Q IV).
Consequently, in quadrants I and III the actuator must be actively supplied with power, while in quadrants II andIV, the actuators can actually supply power to the system.

Figure 5. Definition of load pressure and load flow

STEAM has operating modes which have different ways of connecting an actuator to the three pressure rails, as
shown in figure 6. The switching valves, which are installed at the piston and rod sides of the cylinders and connect
to the three pressure rails, generate nine modes leading to nine different cylinder forces.

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Figure 6. System operating modes

Figure 7(a) describes all nine modes labelled in load quadrants of the hydraulic cylinder for STEAM. For example,
the first label of MP/HP shows that the medium pressure is connected to the piston side and the second label ofthat indicates that the high pressure is connected to the rod side. The position of the modes depends on the HP,
MP and TP pressure levels as well as the piston area ratio 𝛼. The lines passing through quadrants one and three
can be used to actively supply an actuator with flow, while the lines passing through quadrants two and four allow
energy to be recovered from the actuator and stored in one of the accumulators. According to switching modes,
when a resistive force occurs such as 𝑂𝑃1 in figure 7(a), mode HP/HP lying directly above the current operating
point is selected so as to minimize throttling losses. When an assistive force occurs such as 𝑂𝑃2 , MP/MP lying
directly below the current operating point is selected. In this way, throttling losses of the CP-System can be
reduced by using the three pressure rails, however the usage of the three pressure rails cannot promise sufficient
total energy efficiency. Normally the higher pressure level of HP must be set to be equal to the maximum operatingpressure of the conventional machines to keep the same performances. This results in large throttling losses due
to expanding pressure differences between respective modes. Thus, high efficiency for the total system is notexpected due to throttling losses.
To overcome this problem, the basic principle of the new system shown in the previous paper [9] has proposed toreduce the pressure level of HP, as can be seen in figure 7(b). By changing HP from 𝑝HP1 to 𝑝HP2 , it is possible
to reduce throttling losses occurring in regions of respective modes. A problem of the system is that the HP cannot
supply to loads requiring higher pressure levels than 𝑝HP2 . In such a case the main pump can supply the fluid tothe loads directly. Additionally, the flow rate to the actuators with quite lower pressure level than the accumulator
pressure level is provided directly from the pumps. For example, if the actuator with a low pressure level is driven
by the accumulator with a middle pressure level, a large differential pressure resulting in relatively large throttlinglosses will occur. Thus, the flow rate with a low pressure level is provided by the pump directly.

(a) STEAM(b) New system

Figure 7. Comparison of regions of operation against HP

This basic principle of the new system was confirmed in the previous paper [9, 11] by a simulation. However, thesimulation results showed effectiveness of the new system for only the levelling cycle in which actuators are notdriven with large power but with small or middle power. According to a cycle including motions with large power,
it is obvious that system efficiency becomes poor when the accumulator provides flow rate to the actuator withlarge power. For example, when the accumulator provide flow rate to the actuator of 10 kW, which is not large
power, with 50 % of efficiency, losses of 5 kW occur. However, even though efficiency is 80 %, when theaccumulator provides flow rate to the actuator of 100 kW, losses of as much as 20 kW occur. In this way, throttling

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losses resulting from driving the actuator with large power by the accumulator, are likely to affect systemefficiency more. In the levelling cycle used in previous discussion, such a case has not appeared, but in the dig
and dump cycle where actuators are driven with large power, this matter becomes more prominent.
In order to overcome this problem, as can be seen in figure 8(a), the authors propose that actuators with large
power are provided by pumps directly. The region of large power driven by the pumps can be determined basedon the boom power during swing boom lift up motion with the soil because this power is the largest actuator
power required in the dig and dump cycle. In this manner, throttling losses can be reduced. In addition, the ICE
can be driven in high efficiency region when pumps provide large power to the actuators directly, as shown infigure 8(b). The hydraulic circuit of the new system can be designed based on analysis for measurement data of
dig and dump cycle and levelling cycle with this design guideline shown in figure 8(a). Therefore, the new system
can be operated in most motions required by operators because their cycles account for most actual operation of
excavators in the real market.

(a) Design guideline(b) Operation of ICE

Figure 8. Design guideline for new hybrid architecture

NEW SYSTEM DESIGN
Figure 9 shows an outline of dig and dump cycle consisting of four motions. In the digging motion, the arm andbucket are pulled to the machines side to dig and the boom is lifted slightly. Next motion which is swing boom
lift up motion , the boom is largely lifted and the upper structure is rotated 90 degree. In the bucket dump motion,
the bucket is operated to release soil in the bucket. The last motion is to return the initial position, so called swing
boom down motion. These figures show the strokes and angle of each actuator. The cycle time is about 12 seconds.

Figure 9. Dig and dump cycle

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Figure 10 shows histograms of load flow rate and load pressure for each actuator of dig and dump cycle inquadrants presented in figure 5. The horizontal axis indicates load flow for each actuator and the vertical axis
shows load pressure. The color bars represents frequency of the actuator motions which means the ratio of total
to the operation point of the highest frequency. Additionally, red region indicates area of large power and highpressure level. In these region, as shown in design guideline of figure 8(a), the actuators should be powered bypumps directly. By analyzing measurement data of dig and dump cycle with the design guideline, the hydrauliccircuit can be designed. In order to judge the magnitude of the actuator’s powers, constant power hyperbolas areadded to the diagram. In figure 10(a), power of the boom during boom up motion in quadrant one appears in highpower region. Thus, boom up motion is powered by the pump, and accumulator also provide flow rate to the boom
so as to compensate flow rate. On the other hand, boom potential power in the boom down motion can berecuperated by the MP accumulator. In figure 10(b), power of the arm extending the piston in quadrant one doesn’t
appear in high power region mainly. In order to reduce switching pump drive and accumulator drive which resultsin complexity for the hydraulic circuit of the new system, the arm extending the piston should be driven by only
the accumulators. The arm contracting the piston in quadrant three is operated in low power region. Therefore,
flow rate to the arm in quadrant three can be provided by accumulators. In figure 10(c), the load pressure levelsof the bucket extending the piston in quadrant one appear in high pressure level and in small power area. Thus,
the bucket with high pressure level is powered by the pump, and the bucket with small power is driven by theaccumulators. According to swing, figure 10(d) shows that swing drive motions in quadrant three are driven inhigh pressure levels. Therefore, the swing is powered by pumps directly, and the swing brake energy can berecuperated by the HP. From the above, table 1 shows which actuators are powered by the pumps or theaccumulators.

(a) Boom(b) Arm(c) Bucket(d) Swing

Figure 10. Data analysis of dig and dump cycle

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Table 1. Flow distribution matrix in dig and dump cycle
Pump

Accumulator

Boom

Drive

Drive/Recuperation

Arm

-

Drive

Bucket

Drive

Drive

Swing

Drive

Recuperation

Figure 11 shows the hydraulic circuit of the new system made by analysis of measurement data of dig and dump
cycle based on the design guideline. The upper area, framed with the dashed line, shows the open center valves,
and the lower area, framed with the dashed line, shows the constant pressure valves. Since the cylinders can beconnected to the pumps and the tank by using the open center valves, the number of constant pressure valves canbe reduced compared to STEAM. The bottom sides of all cylinders can be connected to HP, and the bottom sides
and rod sides of all cylinders can be connected to MP.

Figure 11. Hydraulic circuit of new system

In figure 12, operations of the new system are presented in the dig and dump cycle. In figure 12(a), about the
digging motion, the pump 1 provides flow rate to the bucket with high pressure levels and the pump 2 provides
flow rate to the boom with high power. The arm with middle power is driven by MP. In swing boom lift up motion,
figure 12(b) shows that the boom and the swing are powered by pump 1 and pump 2, respectively. Furthermore,
the flow rate to the boom is provided by HP so as to compensate flow rate to the boom. When the rotational speedof the swing motor decreases, HP can recuperate swing brake energy. In figure 12(c), according to the bucket
dump motion, the bucket with small power is powered by MP, and at the same time MP and HP can be charged
by the pumps. As can be seen in figure 12(d), in swing boom down motion, pump 2 provides flow rate to theswing and pump 1 charges accumulators, and also, MP provides flow rate to the arm and the bucket. The boom
potential energy and swing brake energy can be recuperated by MP and HP, respectively.

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The 11th JFPS International Symposium on Fluid Power 2020(a) Digging motion(b) Swing boom lift up motion(c) Bucket dump motion(d) Swing boom down motion

Figure 12. Operations of the new system for dig and dump cycle

SIMULATION MODEL
Figure 13 depicts the simulation model. The input signal and control logic are modelled using Simulink, thehydraulic circuit is simulated by AMESim, and a multi-body dynamic model is programmed in Simulation X. All
three programs are run in co-simulation. Measurement data were used as input signals.

Figure 13. Simulation model

For the new hybrid system, considering parameters of the accumulator is crucial since they affect the systemefficiency, initial cost and ease of layout for a real machine. For that, parameter study of accumulator for the8

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accumulator volume and operating pressure is conducted. Table 2 shows parameters of the accumulators.
According to parameters of MP, the same parameters shown in previous paper [11] are used in this paper.
Parameters of operating pressure levels can be decided based on the pressure levels of actuator which should bepowered by MP. For example, about the operating pressure of MP, 150 bar, 160 bar, and 170 bar can be chosenbecause the maximum pressure level of the actuators is approximately 130 bar. Therefore, a little higher pressurelevels than 130 bar are selected for operating pressure levels of MP. According to the volume, if the volume of
the accumulator is too small, pressure levels of accumulator will decrease, and moreover, fluid of the accumulators
will run out. Thus, 100 L for MP is selected to keep the pressure levels of the accumulator higher than actuatorpressure level which is 130 bar. According to the volume of 300 L, it can almost be used as constant pressure in
the simulation. The accumulator volume 200 L is set as a middle volume between the small volume of 100 L andlarge volume of 300 L. On the other hand, according to parameters of HP, the accumulator volumes are same for
MP. Furthermore, the operating pressure of HP are set to 230 bar, 240 bar and 250 bar because the maximum
pressure levels of the boom are approximately 220 bar when the boom is powered by HP in dig and dump cycle.
Table 2. Parameters of volumes and pressure levels for accumulators
Operating pressurelevels [bar]Accumulator
volumes [L]MP

150, 160, 170

100, 200, 300

HP230, 240, 250

100, 200, 300

Table 3(a) shows nine patterns of parameters’ combination about MP from No 1 to No 9. In these patterns, HP is
set to 230 bar and 100 L. On the other hand, table 3(b) shows nine patterns of parameters’ combination about HP
from No 1 to No 9. In these patterns, MP is set to 150 bar and 100 L.
Table 3. Parameter combinations of volumes and pressure levels(a) MP
No
1
2
3
4
5
6
7
8
9

Operating pressurelevels [bar]
150
150
150
160
160
160
170
170
170

(b) HP
Accumulator
volumes [L]
100
200
300
100
200
300
100
200
300

No
1
2
3
4
5
6
7
8
9

Operating pressurelevels [bar]
230
230
230
240
240
240
250
250
250

Accumulator
volumes [L]
100
200
300
100
200
300
100
200
300

SIMULATION RESULTFirst, the simulation is conducted setting MP to 150 bar and 100L, and HP to 230 bar and 100L. Verifying
simulation quality, results of the stroke of cylinders and angle of hydraulic motor are presented in figure 14. Theyshow boom, arm, bucket strokes and swing angle, respectively. The black lines indicate measurement data of the
TPOC-System, and the pink lines indicate the simulation results of the new system. The simulation results match
stroke and angle course quite well.

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Figure 14. Simulation result for stroke and angle

Figure 15 indicates simulation results of cylinder forces and torque of hydraulic motor. They show boom, arm,
bucket forces and swing torque, respectively. The simulation results of the force and torque are in agreement withmeasurement data of TPOC-System. But approximately 5.3 second, swing torque are a little bit different. The
reason is that inertia of soil in the bucket dose not completely correspond to measurement data. However, as
shown in figure 14, the simulation results of the swing angle is same behavior to measurement data, and also, thesimulation results of the swing torque are almost same. Therefore, to compare system efficiency, this simulation
result is useful sufficiently.

Figure 15. Simulation result for force and torque10

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Figure 16 depicts different power courses over time during dig and dump cycle for simulation results only. In
figure 16, from above, power of pump 1, pump 2, HP and MP are presented. During digging motion, pump 1
provides flow rate to the bucket and pump 2 provides flow rate to the boom. The arm is driven by MP. In swing
boom lift up motion, boom and swing are powered by pump 1 and pump 2, respectively. In addition, the boom isdriven by HP. According to the bucket dump motion, the bucket is powered by MP, and at the same time MP ischarged by two pumps. In swing boom down motion, pump 2 provides flow rate to the swing and pump 1 charges
MP and respectively, MP provides flow rate to the arm and the bucket. At the same time, the boom potential
energy is recuperated by MP. Since HP recuperates the swing brake energy, it is not necessary to charge HP by
the pumps. It is recognized that the pump and the accumulator are powers distributed to cylinders as shown intable 1.

Figure 16. Simulation result for power

In figure 17, a sankey diagram is represented to compare the system efficiency of the TPOC-System with the new
hybrid system. The diagram indicates how much energy of diesel fuel was used for driving actuators and how
much dissipated as heat losses in the machine. Energy of fuel consumption 𝐸Diesel can be calculated by ICE’s
efficiency map and ICE’s shaft energy which can be derived based on pump torque, pump rotational speed, pumpefficiency and auxiliary energy. Moreover, a difference between energy of fuel consumption 𝐸Diesel and ICE’s
shaft energy is calculated as ICE losses 𝐸loss.Eng . Auxiliary energy 𝐸aux and pump losses 𝐸loss.pump can be
described by a difference between ICE’s shaft energy and pump energy which is calculated based on pumppressure level and pump flow rate. Actuator energy 𝐸act and recoverable energy 𝐸recoverable can be derived byload pressure and load flow in quadrants one and three of figure 5 and in quadrants two and four, respectively. A
difference between pump energy and sum of 𝐸act and 𝐸recoverable is described as throttling losses in the valvesand hoses 𝐸loss.valve . Recovered energy by the accumulators 𝐸recovered can be calculated based on accumulator
flow from the actuators and accumulator pressure level. A difference between recoverable energy 𝐸recoverable andrecovered energy 𝐸recovered is described as recovery losses 𝐸loss.recovery . According to the TPOC-System infigure 17(a), 63.6 % of diesel fuel is dissipated as ICE losses 𝐸loss.Eng , and also 5.7 % are auxiliary energy 𝐸aux
and pump losses 𝐸loss.pump . Moreover, 12.0 % are throttling losses in the valves and hoses 𝐸loss.valve . The TPOCSystem can not recuperate actuator energy, and therefore recoverable energy 𝐸recoverable is also dissipated as
losses. Thus, 11.1 % of diesel fuel energy is used for driving actuators. On the other hand, for the new system infigure 17(b), 14.6 % of diesel fuel energy could be used in order to power actuators. Reasons for the system

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efficiency improvement are the efficient operation of the ICE and recuperation of boom potential energy andswing brake energy. The new system shows potential of 24 % less fuel compared to TPOC-System.

(a ) TPOC-System(b) New system
Figure 17. Sankey diagram for system efficiency

Figure 18(a) presents the simulation results of the accumulator pressure levels for parameters of MP which are
No 1, No 4 and No 7 in table 3(a). The horizontal axis indicates time, and the vertical axis shows accumulator
pressure levels. All pressure levels which are 150 bar, 160 bar and 170 bar decrease with almost same vales andthen increase due to charging the accumulators by the pumps and recoverable energy. However, according to 160
bar and 170 bar, at the end of the cycle, the accumulator pressure levels are lower than the start of accumulator
pressure levels, especially 170 bar. The reason is that throttling losses resulting from driving the actuator by 170
bar accumulator is lager than using 150 bar accumulator because pressure difference is larger. This causes increaseof energy to charge the accumulator. However, the pump energy to charge the accumulator is not enough in lower
rotational speed of the ICE. Therefore, at the end of cycle, the accumulator pressure level is lower. Figure 18(b)indicates simulation results of pressure levels for No 1, 2, and 3 in table 3(a). In figure 18(b), the pressure changeof the accumulator for 100 L increases, and the pressure change of the accumulator for 300 L decreases. These
characteristics will be discussed later.

(a) MP: 150 bar, 160 bar and 170 bar(b) MP: 100 L, 200 L and 300 L

Figure 18. Simulation result for MP accumulator pressure level

Figure 19(a) depicts the simulation results of the accumulator pressure levels for parameters of HP which are No
1, No 4 and No 7 in table 3(b). As with the simulation result of MP, all pressure levels decrease with almost same
vales, and then increase. Furthermore, the start and end of the pressure levels are almost same because swing

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brake energy is enough to charge the accumulator. Figure 19(b) presents simulation results of pressure levels for
No 1, 2, and 3 in table 3(b). In figure 19(b), the pressure change of the accumulator for 100 L increases, and thepressure change of the accumulator for 300 L decreases.

(a) HP: 230 bar, 240 bar and 250 bar(b) HP: 100 L, 200 L and 300 L

Figure 19. Simulation result for HP accumulator pressure level

The pressure change of the accumulator against the change in volume can be explained with figure 20. Figure 20
shows the different operating states of the accumulator. 𝑝0 is the pre-charge gas pressure, and 𝑉0 is the maximum
gas volume at 𝑝0 . 𝑝1 is the lower operating pressure, and 𝑝2 is the higher operating pressure. The change in stateof the accumulator can be shown by Eq. (1).

𝑝0 ∙ V k0 = p1 ∙ V k1 = 𝑝2 ∙ V k2 = const()

A cycle time of the dig and dump cycle which is about 12 second. This means that the liquid is removed quickly
from the accumulator during dig and dump cycle. The change of state of the accumulator, thus, is defined asadiabatic changes of state. The adiabatic index k is set to 1.4. From the above, Eq. (2) can be derived from Eq (1).
1 k

1 k

2

1

∆𝑝 = 𝑝0 ∙V k0 [ (V ) - (V ) ]()When the volume of the accumulator 𝑉0 is sufficiently large (300 L) and changed from 𝑉2 to 𝑉1 , the pressurechange ∆𝑝 becomes small because the change in volume ∆𝑉 is quite small compared to the accumulator volume
𝑉0 . On the other hand, in a case where the volume of the accumulator 𝑉0 is not large (100 L) and it changes from
𝑉2 to 𝑉1 , the pressure change ∆𝑝 become larger than that of the accumulator with a large volume. If theaccumulator volume 𝑉0 is large, the pressure change ∆𝑝 will be small because the volume change ∆𝑉 is relatively
small compared to the large accumulator volume 𝑉0 . If the accumulator volume 𝑉0 is small, the pressure change∆𝑝 will be large because the volume change ∆𝑉 is relatively large compared to the small accumulator volume 𝑉0 .
From the above scenarios, the simulation result with 100 L has the largest pressure change.

Figure 20. Parameters of an accumulator

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According to MP parameters, to compare the system efficiency at pressures of 150 bar and 170 bar, a Sankey
diagram of the 170 bar pressure level is presented in figure 21(a). Compared to the 150 bar accumulator pressure
shown in figure 17(b), throttling losses in the valves at 170 bar increase because the differential pressure between
the accumulator and actuators is larger. Therefore, the system efficiency of 170 bar become worse. The Sankey
diagram of the 300 L is shown in figure 21(b). Compared to the 100 L accumulator shown in figure 17(b), systemefficiency is almost equal because throttling losses are same as well.

(a) MP: 170 bar (No 7 in table 3(a))(b) MP: 300 L (No 3 in table 3(a))
Figure 21. Sankey diagram for MP parameters

According to HP parameters, to compare the system efficiency about pressures and volumes, Sankey diagrams ofthe 250 bar pressure level with 100 L and 300 L with 230 bar are shown in figure 22(a) and figure 22(b),respectively. Compared to the simulation results in figure 17(b), the system efficiency of 250 bar and 300 L isalmost same because throttling losses are in agreement with the simulation results in figure 17(b).

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The 11th JFPS International Symposium on Fluid Power 2020(a) HP: 250 bar (No 7 in table 3(b))(b) HP: 300 L (No 3 in table 3(b))
Figure 22. Sankey diagram for HP parameters

Figure 23(a) shows the influence of changing parameters for MP from No. 1 to 9 in table 3(a), in terms of energy
savings, reduction in fuel consumption against the volumes, and the pressure level. These values compare the fuelconsumption of the new system with that of the reference TPOC-System. If the volume of the accumulator is set
to a smaller volume, the reduction in fuel consumption is slightly improved. On the other hand, changing thepressure level has a greater effect on the system efficiency of the new hybrid system than changing the volume of
the accumulator. Thus, from the perspective of improving energy saving, it is quite important that the pressurelevel of the accumulator is set as low as possible.
Figure 23(b) shows the influence of changing parameters for HP from No. 1 to 9 in table 3(b), in terms of energy
savings, reduction in fuel consumption against the volumes, and the pressure levels. Parameters of HP do notlargely affect the system efficiency because discharged and charged power of HP is smaller than MP as shown infigure 16.
100 L

200 L

100 L

300 L

Reduction in fuel consumption [%]Reduction in fuel consumption [%]

24
23
22
21
20
19
18
150

160
Operating pressure levels [bar](a) MP

200 L

300 L

25

25

170

24
23
22
21
20
19
18
230

240
Operating pressure levels [bar](b) HP

250

Figure 23. Reduction in fuel consumption for all simulation results

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ACKNOWLEDGMENTS
This research work was supported by Hitachi Construction Machinery Co., Ltd.

CONCLUSION
In order to develop the hydraulic circuit of the new hybrid system, the design guideline was proposed from theperspective of energy efficiency and a simple configuration for dig and dump cycle. Moreover, the hydrauliccircuit for the new system was designed based on measurement data of the dig and dump cycle with the design
guideline. As a result, in simulation, the new system consumes 24 % less fuel than a conventional system which
is TPOC-System. Furthermore, the parameters, which are the accumulator volume and pressure level, arediscussed in this paper because they have a significant effect on fuel energy savings and initial cost. Throughsimulation, it can be recognized that setting the MP accumulator pressure level as low as possible is quite effective
for energy saving. Furthermore, parameters of HP do not largely affect the system efficiency because discharged
and charged power of HP is smaller than MP.

REFERENCES
1. EGAWA, H., KAWASAKI, H., Study of Electro-Hydraulic Energy Saving System (EHESS) for Construction
Machinery,12th Scandinavian International Conference on Fluid Power (12. SCIFP), Tampere, Finland, 2011.
2. BUSQUETS, E., IVANTYSYNOVA, M., Toward Supervisory-Level Control for the Energy Consumption and
Performance Optimization of Displacement-Controlled Hydraulic Hybrid Machines, 10th International Fluid Power
Conference (10. IFK), Dresden, Germany, 2016.
3. SIEBERT, J., GEIMER, M., Reduction of System Inherent Pressure Losses at Pressure Compensators of Hydraulic Load
Sensing Systems, 10th International Fluid Power Conference (10. IFK), Dresden, Germany, 2016.
4. DENGLER, P., GEIMER, M., BAUM, H., SCHUSTER, G., WESSING, C., Efficiency Improvement of a ConstantPressure System using an Intermediate Pressure Line, 8th International Fluid Power Conference (8. IFK), Dresden,
Germany, 2012.
5. PÖTTKER, A., Mechanic – hydraulic co-simulation on mining excavators, 5th International Fluid Power Conference (5.
IFK), Aachen, Germany, 2006.
6. NAKAMURA, T., NAKAMURA, K., OKANO, Y., ISHIKAWA, K., SATO, K., AZUMA, H., KANEHAMA, M.,
KAJITA, Y., Hydraulic Drive Device for Working Machine, Japan Patent, 2012, WO 2012/157705 A1, JP 2012-241803A.
7. VUKOVIC, M., MURRENHOFF, H., STEAM – A Mobile Hydraulic System with Engine Integration, ASME/BATH
2013 Symposium on Fluid Power & Motion Control, Sarasota, Florida, USA, 2013.
8. VUKOVIC, M., LEIFELD, R., MURRENHOFF, H., STEAM – a hydraulic hybrid architecture for excavators, 10th
International Fluid Power Conference (10. IFK), Dresden, Germany, 2016.
9. HIJIKATA, S., WEISHAAR, P., SUGIMURA, K., SCHMITZ, K., MURRENHOFF, H., A hydraulic hybrid architecture
combining an open center with a constant pressure system for excavators, 11th International Fluid Power Conference (11.
IFK), Aachen, Germany, 2018.
10. HIJIKATA, S., WEISHAAR, P., LEIFELD, R., SCHMITZ, K., Efficiency for a hydraulic hybrid architecture of excavators,
MM science journal, 2018, DOI:10.17973/MMSJ.2018_10_201831.
11. HIJIKATA, S., ITO, K., MURRENHOFF, H., Investigation of Accumulator Parameters for a Novel Hybrid Architecture,
J. Robot. and Mechatronics, 2020, Vol.32 No.5, pp. 876-884.

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Research Regarding the Energy Saving Conditions of the Air Blow
for Removing and Drying Out Water
Daisuke KURAKAMI*, Keiichirou KOGA*, Gohei HARIMOTO**

Technology Research Division, SMC Corporation
4-2-2, Kinunodai, Tsukubamirai-shi, Ibaraki-ken, 300-2493 Japan(E-mail: kurakami.daisuke@smcjpn.co.jp)

Abstract. One example of the use of air blow is drying workpieces after cleaning. The surface shape of some
workpieces is complicated and there are various holes and dents. Hence it can’t be dried well, even if a large amountof compressed air is used. If may be the cause of the increased power consumption of the air compressor. Therefore,
we have been conducting research on air blow with the aim of establishing a method for selecting appropriate air blow
conditions. In this research, we conducted a flat plate drying experiment using an air blow using a convergent nozzle,
and summarized the relationship between blow time and the dried range. From the results, we have created a diagramthat can select the air blow conditions in which the air consumption is the least according to the required dried range,
and we will introduce it.
Keywords: Air blow, Convergent nozzle, Drying, Energy saving

INTRODUCTION
Air blow is used in various applications and adopted in most of the factories. It is considered that approximately
50% of the compressed air used in the factory is occupied by air blow. Therefore, the improvement of the existing
air blow equipment is cost-effective, and the user’s interest and awareness of it is high [1]. If a discharge pressureof an air compressor can be reduced by efficient air blow, it will lead to reduction of power consumption andenergy saving.
One example of the use of air blow is drying workpieces after cleaning. The surface shape of some workpieces iscomplicated and there are various holes and dents. Hence it can’t be dried well, even if a large amount ofcompressed air is used in the drying process after cleaning. This is the problem for many users. Therefore, it is an
issue to remove water and improve drying efficiency by air blow. Based on the above, we have been conducting
research on air blow with the aim of establishing a method for presenting appropriate air blow conditions. Here,
air blow conditions include nozzle diameter, nozzle inlet pressure, and distance from a workpiece.
In this paper, we will introduce the diagrams to select appropriate blow conditions when drying a flat plate.

EXPERIMENT DRYING FLAT PLATE
Experimental device
Fig. 1 shows the diagram of experimental device. The nozzle inlet pressure (hereinafter referred to as “P2 [MPa]”)
is regulated by the pressure regulator. The set pressure of the pressure regulator (often equal to the dischargepressure of the air compressor) is P1 [MPa]. P2 is measured by the pressure sensor. The flow rate (hereinafterreferred to as “Q [L/min(ANR)]”) consumed during air blow is measured by the flow sensor. The position of thenozzle is adjusted by electric actuators. The size of the aluminum plate mounted under the nozzle is 500 mm
square.
Before conducting the experiment, some preliminary experiments were conducted and the following was
confirmed. Firstly, the nozzle-plate distance has little effect on the dried diameter in the range of 50 mm to 160
mm. Secondly, when P2 is less than 0.1 MPa, that is, in the case of subsonic flow, it is necessary to increase boththe blow time and the air consumption in order to obtain the same drying effect as the choked flow. Consequently,
in this experiment, the distance between the nozzle and the plate is 120 mm and P2 shall be 0.1 MPa or more.
Tab. 1 shows experimental conditions. In this experiment, a nozzle called a convergent nozzle is used. Fig. 2
shows the cross-sectional shape of a convergent nozzle.

1

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Set pressure: P1

Flow rate: Q

Nozzle inlet pressure:P2 (0.1 – 0.6 MPa)

Dried diameter

Nozzle - Plate distance:
120 mmWater

Figure 1. Experimental device diagram

Nozzle diameter:d [mm]
1
2
3
4
6

Table 1. Experimental conditions
Nozzle inlet pressure: P2 [MPa]
0.1
0.2
0.3
0.4
0.5

0.6

○
○
○
○
○

○
○
○
-

○
○
○
○
○

○
○
○
○
-

○
○
○
○
-

○
○
○
-

Nozzle diameter: d [mm]Figure 2. Cross-sectional shape of convergent nozzle

Experimental method
First, spread water evenly on the flat plate. Next, the surface is dried by air blowing. The time keeping air blow is
60 seconds. This situation is recorded as a video with a digital camera. From the video, the image after a certain
period of time from the start of air blow is acquired. Use image processing software to measure the dried diameterfrom the acquired image.
Fig. 3 shows an example of an image taken from a video. For each experimental condition, record the relationshipbetween the time from the start of blowing (hereinafter referred to as “blow time: t [s]”) and the dried diameter(hereinafter referred to as “D [mm]”).

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3-port valve

Pressure sensor

NozzleDried diameter: D

Aluminum plate

Figure 3. Dry image example

Experimental results
As an example of the experimental results, Fig. 4 shows the relationship between D and t when d = 1 mm,
P2 = 0.2 MPa. D significantly changes immediately after starting the air blow, however the rate of the change ofit reduces over time.

Dried diameter: D [mm]

300
250
200
150
100
50
0
0

10

20

30
40
Blow time: t [s]5060

Figure 4. Relationship between dried diameter and blow time(d =1 mm, P2 = 0.2 MPa)

CREATION OF THE NOZZLE SELECTION DIAGRAMBASED ON THE ENERGY SAVING CONDITIONS
Relationship between the maximum dried diameter and flow rate
Fig. 5 shows a plot of the experimental values for the relationship between Q and the maximum dried diameter(hereinafter referred to as “Dmax [mm]”). The shape of the markers is changed for each d. Here, Dmax means the Dafter the end of the air blow. From the graph, it is estimated that when using a convergent nozzle, Dmax isdetermined by Q, regardless of d or P2. The relationship between Dmax and Q can be approximated by Equation 1.
The solid line in Fig. 5 shows the value calculated by Equation 1.

D max = 125 × Q 0 . 245(1)

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Maximum dried diameter: Dmax [mm]The 11th JFPS International Symposium on Fluid Power 2020

700
600
500
mm
dd==11mmdΦ2= 2 mmdΦ3= 3 mmdΦ4= 4 mmdΦ6= 6 mm
Eq. 1

400
300
200
100
0
0

200

400
600
800
Flow rate: Q [L/min(ANR)]

1000

Figure 5. Relationship between maximum dried diameter and flow rate

Relationship between the dried diameter and blow time
Based on the experimental results under all the conditions shown in Tab. 1, the relationship between D and t canbe expressed by an equation such as the exponential function shown in Equation 2. D0 [mm] represents the initialD, and it is calculated by Equation 3.
t−
D = D 0 + (D max − D 0 ) 1 − e 12 D 0 = 12 d + 9

(2)
(3)

Fig. 6 shows the relationship between t and D when d = 2 mm, P2 = 0.2 or 0.5 MPa. The experimental values are
represented by markers, and the values calculated by Equation 2 are represented by solid lines.

Dried diameter: D [mm]

500
400
300

0.2MPaPP2
MPa
2 ==0.2
P0.5MPa
2 = 0.5 MPa

200

Eq. 2

100
0
0

10

20

30
40
Blow time: t [s]5060

Figure 6. Relationship between dried diameter and blow time(d = 2 mm, P2 = 0.2 or 0.5 MPa)

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Air-saving blow condition selection diagram
Using Equation 1 to 3, t on which D is reached to Φ200 to Φ600 and the air consumption (hereinafter referred toas “q [L(ANR)]”) at that time are calculated. Fig. 7 shows the result. Comparing the air blow conditions when D
is the same, the following tendencies can be read. When d is the same, the lower P2 is, the less q is, however the
longer required t is. Similarly, when t is the same, the larger d is, the lower P2 is and the less q is.
As a selection example, consider the case of drying a range with a diameter of Φ500. The marker on Fig. 7 withthe least q in markers indicating air blow conditions in which D is reached to Φ500 is circled in red. The conditions
indicated by this marker are d = Φ3, P2 = 0.5 MPa, and t = 25 s. In this way, the air blow conditions with the leastq can be selected according to the desired D. The least q means that the best condition for Energy saving.

Air consumption: q [L(ANR)]

600
500

0.1MPaDried diameter

0.2MPaΦ600

Nozzlediameter
Φ60.3MPaΦ4

0.4MPa400

Φ500

0.5MPa0.6MPa300

Optimum
conditon

Φ400

Φ3

200
192

100

Φ300
Φ2

Φ200
Φ1

0
0

5

10

15
20
25
Required blow time: t [s]

30

35

40

Figure 7. Diagram for selecting the nozzle for drying the flat plate

Energy saving nozzle selection diagram
according to set pressure of compressor
In the selection diagram shown in Fig. 7, it is possible to select the air blow conditions in which q is the least,
according to the desired D. Incidentally, a discharge pressure of an air compressor varies from factory to factory.
Therefore, in order to set to the P2 selected by using Fig.7, the discharge pressure must be reduced by a pressure
regulator. However, if the pressure adjustment range is large, energy loss will occur. Thus, it is efficient that P2 is
as close as possible to a discharge pressure of an air compressor. From the above, it is better to compare q for eachd under the same P2. Consequently, a nozzle selection diagram for each P2 has been created so that the optimum
d can be easily selected for each P2. Fig. 8 (a) to (d) respectively show the relationship between D and q for eachd when P2 is 0.3 to 0.6 MPa.
Using these diagrams, it is possible to choose d with the least q, depending on P2 and the desired D. For example,
assume that P2 = 0.5 MPa. In this case, see Fig. 8 (c). From this diagram, d with the least q according to the desired
D is determined. Tab. 2 shows the correspondence table between D and the appropriate d when P2 = 0.5 MPa. In
this way, since the appropriate d differs depending on the pressure used and the desired D, it is necessary to use
different nozzles according to the blow conditions.

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Φ2 Φ3

500

Φ1

400

Φ6

300
Φ4

200
100
0
0

100 200 300 400 500 600
Dried diameter: D [mm]Air comsumption: q [L]Air comsumption: q [L]The 11th JFPS International Symposium on Fluid Power 2020

500
400
300

Φ4

200
Φ6

100

Φ2

0
0

Φ4

400
300

Φ3

200
Φ6
Φ2

Φ1

0
0

100 200 300 400 500 600
Dried diameter: D [mm](b) P2 = 0.4 MPa

100 200 300 400 500 600
Dried diameter: D [mm]Air comsumption: q [L]Air comsumption: q [L](a) P2 = 0.3 MPa

100

Φ3

Φ1

400

Φ1

300
Φ4

200

Φ3

Φ6

100

Φ2

0
0

(c) P2 = 0.5 MPa

100 200 300 400 500 600
Dried diameter: D [mm](d) P2 = 0.6 MPa

Figure 8. Relationship between dried diameter and air consumption(P2 = 0.3 to 0.6 MPa)
Table 2. Correspondence tableof dried diameter and nozzle diameter(P2 = 0.5 MPa)
DesiredAppropriate
Nozzle diameter [mm]dried diameter [mm]- 320
320 - 460
460 - 550
550 -

Φ1
Φ2
Φ3
Φ4

CONCLUSION
Based on the experimental results, equations have been established to calculate the dried diameter from flow rate,
nozzle diameter, and blow time when drying a flat plate with a convergent nozzle.
Using the established equations, two types of diagrams were created to select the appropriate air blow conditions
according to the desired dried diameter.
In the first type of diagram, the air blow condition with the least air consumption can be selected from the allpressure conditions. This type of diagram can include blow time as a selection condition.
In the second type of diagram, the nozzle with the least air consumption can be selected for each pressure. If thepressure to be used is decided, it is suitable to use this type of diagram.

REFERENCES
1. ONEYAMA. N., Energy Saving for Pneumatic System, Energy Conservation Center, 2003, pp.11-12, 97, 195-196.

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Energy Regeneration and Reuse of Excavator Boom System With
Hydraulic Constantly Variable Powertrain
Tri Cuong DO* and Kyoung Kwan AHN***

Graduate School of Mechanical and Automotive Engineering, University of Ulsan, Ulsan, Korea**
School of Mechanical and Automotive Engineering, University of Ulsan, Ulsan, Korea(E-mail: kkahn@ulsan.ac.kr)

Abstract. A new design of an energy recovery system is presented in this paper for regenerating and reusing
potential energy in a boom cylinder. The proposed system integrates an electrical hydraulic continually variable
powertrain which the main pump is supplied power from both engine and an electric motor/generator. Duringmoving down process, the potential hydraulic energy of boom cylinder is stored in a battery as electricity through ahydraulic motor and an electric motor/generator. Then, during acceleration process, the stored electrical energy canbe reused to the electric motor/generator (motor mode) to support the engine. To demonstrate the effectiveness of the
proposed system, a real test bench is built. The experiment results indicated that the energy regeneration efficiency
could reach up to 37%. Besides, the energy consumption was also reduced.
Keywords: Boom system, constantly variable powertrain, energy recovery system.

INTRODUCTION
In recent years, with the development of industry, pollution and global warming have become extremely serious
problems. One of the most popular and widely used equipment is the hydraulic excavator because of its benefits
such as moving heavy objects, dig, and demolition. However, like other construction machinery, hydraulicexcavators consume a lot of fossil fuels and release harmful gases into the environment [1-4]. To solve thisproblem, there are two main directions that are to reduce energy consumption and reuse the potential energy inthe excavator.
To reduce energy consumption, conventional hydraulic excavators (HE) with internal combustion engines(ICEs) are gradually being replaced by hybrid HE that combines the ICE with an electric motor or electric HE
using electric motor and battery. The first structure of hybrid HE is proposed by Kanezawa [5]. After that, a fullscale commercial hybrid HE namely PC200-8 is released from Komatsu [6]. In this hybrid HE, the potential
energy in the swing system is recovered through a generator and stored in the battery. The results showed thatthe energy consumption was reduced 25% compared with conventional HE. Lei Ge proposed an electric HE
configuration [7]. In this research, a speed variable electric motor was used as the main power source to drive avariable displacement pump. Besides, a control strategy was designed to adjust the pump displacement andmotor speed to ensure these components working in highly efficient areas. The experiment results indicated thatthe energy consumption reduced from 28.5 to 33%. However, the electric HE still has limitations such as low
durability and not suitable for dusty environments.
In HE, there are some places where the energy can be recovered such as in the boom cylinder during the moving
down or flow rate of the swing system during deceleration process. In which, potential energy in boom cylinder
accounts for 51% of the total recoverable energy [8]. Therefore, energy recovery systems (ERSs) for HE focuseson the recovery of energy in boom cylinder. Yu et al. [9] placed a hydraulic motor at the outlet port of borechamber to reduce the energy losses. In addition, a flow control valve was integrated to regulate the flow rate
through the hydraulic motor. Hence, the hydraulic motor’s working point is ensured in the high-efficiency area.
The experiment result showed that the energy regeneration efficiency can achieve 57,4%. Based on thisresearch, Do et al. [10] proposed a ERS which can not only recover but also reuse the energy in the boom
cylinder system. The proposed system used energy conversion components which combined a hydraulicpump/motor and an electric motor/generator. The boom hydraulic energy is converted to electrical energy andstored in the battery. Then, the stored energy can be reused to the electric motor/generator (motor mode) in thenext cycles. However, this system required large device sizes to perform the above functions. This leads toincreased manufacturing costs and payback times. Besides, when applying this ERS to the hybrid HE, thesystem has to integrate an additional electric motor/generator to recover energy, leading to an increase in thecost of the hybrid HE.
Based on the above analysis, this paper proposes a new ERS using an electrical hydraulic continually variable
powertrain which contains an engine and an electric motor/generator to drive the main pump. The proposed1

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system can study not only the problem of reducing energy consumption at the power sources but also the
problem of energy recovery in the boom cylinder which have often been studied independently in previous
systems. During the moving down process, the potential energy of the boom cylinder is recovered through ahydraulic motor and the electric motor/generator (generator mode) and stored in the battery. Therefore, it ispossible to overcome the weakness of previous research [10] with only one electric motor/generator in the
powertrain of the hybrid configuration for energy recovery and energy reuse. During the lifting up, the main
pump is driven by the ICE and the electric motor/generator (motor mode) which is supplied power from thebattery. With the proposed structure, the system is not only recover but also reuse the energy in the boom
cylinder system. To prove the effectiveness of the proposed system, a real test bench is built. Several
experiments are conducted with the different working modes of the system. The experiment results indicate thatthe energy regeneration efficiency could achieve 40%. Besides, energy consumption is reduced.
The rest of this paper is organized as follows: Section 2 presents the structure and working principle of the
proposed system. The experiment results are presented in section 3. Conclusions are given in the final section.

SYSTEM DESIGN
The proposed system consists of a boom cylinder, an electric motor which is used to emulate the ICE, an electricmotor/generator, a variable displacement pump, two proportional control valve 4/3, and some auxiliary valves
for safety as shown in Fig. 1. In details, the motor (1) and the electric motor/generator (5) drive the main pump(10) via a planetary gear (3). The rotation speed from the electric motors is connected or disconnected by clutch
and brake devices (2, 4, 9). Besides, the double clutch (8) is used to switch the connection between the main
pump (10), hydraulic motor (14) and the planetary gear (3). Based on the structure, the proposed system canoperate with three main modes: normal working, energy regeneration and energy reuse.
12

Boom cylinder
Load

Energy regeneration part

13

11

V2

V1

14

10

9
3
1

2

Motor

Clutch/Break

Clutch/Break

Clutch/
Break

8

4

DoubleClutch5Electric Motor/
Generator

6

7

Inverter/Converter

Battery

Figure 1. The proposed boom ERS.

In the proposed system, the planetary gear (3) is one of the key components for transferring the power fromelectric motors (1, 5) to the main pump (10). The planetary gear (3) contains sun, carrier and ring gears asshown in Fig. 2. The electric motor (1) is connected to the sun gear, the electric motor/generator (5) and themain pump (10) are connected with the carrier gear and the ring gear, respectively.

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Figure 2. Structure of planetary gear.
Boom cylinder
Load
Energy regeneration partV2

ClutchRing gear

Electric
Motor

Clutch

V1

Sun
Gear

Break

DoubleClutchElectric Motor/
Generator

Inverter/Converter

Battery

Carrier Gear

Figure 3. Working principle of normal mode.
Boom cylinder
Load
Energy regeneration part

V2

Ring gear

Electric
Motor

Break

Sun
Gear

V1

ClutchDoubleClutchClutchElectric Motor/
Generator

Inverter/Converter

Battery

Carrier Gear

Figure 4. Boom cylinder moves down in the energy regeneration mode.

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Boom cylinder
Load
Energy regeneration part

V2

V1

ClutchDoubleClutchClutchGenerator

Ring gear

Electric
Motor

ClutchSun
Gear

Inverter/Converter

Battery

Carrier Gear

Figure 5. Boom cylinder moves up in the hybrid mode.
Boom cylinder
Load
Energy regeneration part

V2

Ring gear

Electric
Motor

Clutch

Sun
Gear

V1

ClutchDoubleClutchClutchElectric Motor/
Generator

Inverter/Converter

Battery

Carrier Gear

Figure 6. Boom cylinder moves up in the energy reuse mode.
In the normal working mode, the main pump (10) is driven by the electric motor (1) and supply the flow rate to
the boom system. The movement of the boom cylinder (12) is controlled by the state of valve V1 (11). When theboom cylinder moves down under the influence of gravity, it can automatically move down without suppling theflow from the power source. The flow rate from the tank can go directly to the rod chamber via valve V1 (11) asshown in Fig. 3.
In the energy regeneration working mode, the boom cylinder (12) moves down by the gravitational force.
Instead of going straight to the tank as in normal working mode, the directional valve V2 (13) is adjusted to theleft position allowing flow from the bore chamber to pass through a hydraulic motor (14) as shown in Fig. 4.
Then, the hydraulic energy is converted to electrical energy through the electric motor/generator (5) and store in
the battery (7).
In the hybrid mode, three clutches/brakes (2, 4, 9) work in the clutch function and connect two power sources toplanetary gear. During this operation, the electric motor (1) supplies the power to both the main pump (10) andelectric motor/generator (5). The cylinder (12) moves up by the flow rate from the main pump (10). Besides, theelectric motor/generator (5) works in generator function and convert the mechanical energy to the electricenergy and store in the battery (7) as shown in Fig. 5.
In the energy reuse mode, the stored energy in the battery (7) is reused and supplied to the electricmotor/generator (5). Then, the electric motor/generator (5) works as motor mode and drives the main pump (10).
The flow rate from the tank is supplied to the boom cylinder (12) and lift the load. The flow rate in the rod
chamber goes to the tank as shown in Fig. 6.

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Table 1. Proposed system configuration.

Components

Parameter
Bore diameter

Value
50

[SI unit]mm

CylinderRod diameter

28

mm

Length

0.75

m

Pump

Displacement

35

cm3/rev

Electrical motor

Power

7.5

kW

Hydraulic motor
Generator/
motor

Displacement

20

cc/rev

Power

5.5

kW

Figure 7. Experiment apparatus of proposed system

EXPERIMENT RESULTS
The experimental setup was built to verify the effectiveness of the proposed system. The detail structure of theexperimental system using electrical hydraulic constantly variable powertrain is presented in Fig. 7. The total
system includes the main parts such as the cylinder, powertrain, hydraulic circuit and control box. The
parameters of the experimental system are shown in Table 1. As presented above, the ICE was emulated by the
7.5kw electric motor. The system is controlled on the computer within real-time Window Target Toolbox of
MATLAB 2013b under a sampling time of 0.01 seconds.

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Pressure (Pa)

Figure 8. Displacement of the boom cylinder. (NM: normal mode, RM: regeneration mode, HM: hybrid mode,
RUM: Reuse mode)

Figure 9. Pressure of the boom cylinder.

Figure 10. Flow rate of the main pump.

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Speed (rpm)The 11th JFPS International Symposium on Fluid Power 2020

Torque (Nm)Figure 11. Speed of electrical motor/generator and ICE.

Figure 12. Torque of electrical motor/generator and ICE.
104

6

5

4

3

2

1

0

0

10

20

30

40

50

60

70

80

Time (s)Figure 13. Speed and torque of electrical motor/generator.

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Fig. 8 presents the displacement of the boom cylinder in three cycles. From the 0s to 10s, the system run in thenormal mode. The boom cylinder took 5 seconds to reach 0.68 m. From the 20s to 30s, the hybrid mode isactivated. During this operation, the ICE supplied the power to both cylinder and electric motor/generator. Fromthe 42s to 58, the system operated in the reuse mode with the main pump driven by the electric motor/generator
without power from the ICE. During the moving down process, the pressurized flow rate from the cylinder isconverted to electric energy through the hydraulic motor and electric generator in energy regeneration mode.
Fig. 9 and Fig. 10 shown the pressure and flow rate of the main pump, respectively.
The speed and torque of the electrical motor/generator and engine were presented in Fig. 11 and 12,
respectively. In the regeneration mode, the electric motor/generator worked in the generator function. Therefore,
the torque and speed were different signs. From the 42s to 58, the speed and torque were the same signs because
the electric motor/generator was working in the motor functions and driving the main pump in the energy reuse
mode. The highest torque and speed of the electric motor/generator were 20 Nm and 320 rpm during threecycles. Fig. 13 shows the energy consumption of the proposed system. From the 30s, the system could workwithout the power from the ICE. Hence, energy consumption could be reduced.
The potential energy in the boom cylinder can be estimated based on the pressure (pc) and flow rate (qc) as
E p   pc qc dt(1)
The regenerated energy can be estimated as shown in Eq. (2).
Er   Tr r dt(2)

where the Er, Tr, and r are the regenerated energy, torque and speed of the electric motor/generator,
respectively. The energy regeneration efficiency of the proposed system can be estimated as follow:
E
(3)
 r
Ep
From the above results, the energy regeneration efficiency of the proposed system was 37%.

CONCLUSION
This paper presented a new energy regeneration system using the electrical hydraulic constantly variable
powertrain. In details, the proposed contained a hydraulic motor that could regenerate the energy from the boom
cylinder and stored in the battery. In addition, the drive system includes an ICE and an electric motor / generator
allowing it to reuse energy during operation. A test rig was developed to prove the effectiveness of the proposedsystem. The results indicated that the system could work well with designed modes. The energy recovery
efficiency of the proposed system could achieve 37%. Besides, the energy consumption was decreased becauseof energy reuse mode.

ACKNOWLEDGMENT
This research was supported by Basic Science Program through the National Research Foundation of Korea(NRF) funded by the Ministry of Science and ICT, South Korea (NRF-2020R1A2B5B03001480).

REFERENCES
1. DO, T. C., TRUONG, H. V. A., DAO, H. V., HO, C. M., TO, X. D., DANG, T. D., and AHN, K. K., Energy
Management Strategy of a PEM Fuel Cell Excavator with a Supercapacitor/Battery Hybrid Power Source, Energies, vol.
12, no. 22, 2019.
2. DANG, T. D., DO, T. C., TRUONG, H. V. A., HO, C. M., DAO, H. V., XIAO, Y. Y., JEONG, E. J. and AHN, K. K.,
Design, Modeling and Analysis of a PEM Fuel Cell Excavator with Supercapacitor/Battery Hybrid Power Source,
Journal of Drive and Control, 16(1), 45-53, 2019.
3. TRUONG, H. V. A., DAO, H. V., DO, T. C., HO, C. M., TO, X. D., DANG, T. D., and AHN, K. K., Mapping Fuzzy
Energy Management Strategy for PEM Fuel Cell–Battery–Supercapacitor Hybrid Excavator, Energies, 13(13), 3387,
2020.
4. DINH, T. X., THUY, L. K., TIEN, N. T., DANG, T. D., HO, C. M., TRUONG, H. V. A., DAO, H. V., DO, T. C., andAHN, K. K., Modeling and Energy Management Strategy in Energetic Macroscopic Representation for a Fuel Cell
Hybrid Electric Vehicle, Journal of Drive and Control, 16(2), 80-90, 2019.
5. KANEZAWA, Y., DAISHO, Y., and KAWAGUCHI, T., Increasing efficiency of construction machine by hybrid
system. Society of Automotive Engineers of Japan. Vol. 100. 2001.
6. INOUE, H., Introduction of PC200-8 Hybrid Hydraulic Excavators. Komatsu technical report 2008, Vol. 54 No.161.
7. GE, L., QUAN, L., ZHANG, X., ZHAO, B., YANG, J., Efficiency improvement and evaluation of electric hydraulicexcavator with speed and displacement variable pump. Energy Conversion and Management 2017, 150, 62-71.

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The 11th JFPS International Symposium on Fluid Power 2020

8. LIN, T., CHEN, Q., REN, H., HUANG, W., CHEN, Q., FU, S., Review of boom potential energy regeneration
technology for hydraulic construction machinery. Renewable and Sustainable Energy Reviews 2017, 79, 358-371.
9. YU, Y. X., AHN, K. K., Optimization of energy regeneration of hybrid hydraulic excavator boom system. Energy
Conversion and Management 2019, 183, 26-34.
10. DO, T. C., NGUYEN, D. G., DANG, T. D.; AHN, K. K., A Boom Energy Regeneration System of Hybrid HydraulicExcavator Using Energy Conversion Components, Actuators, 2021.

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Efficient Closed Pump Controlled Hydraulic-Gas Balanced Energy
Recovery Driving Method for Hydraulic Excavator Boom
Lianpeng XIA*, Long QUAN*, Hongjuan ZHANG**, Yunxiao HAO*, Lei GE**

College of Mechanical and Vehicle Engineering, Taiyuan University of Technology
79 Yingzexi Street, Taiyuan, Shanxi, 030024 China(E-mail: Quanlong@ tyut.edu.cn)**
College of Electrical and Power Engineering, Taiyuan University of Technology
79 Yingzexi Street, Taiyuan, Shanxi, 030024 China

Abstract. To reduce the energy consumption and pollution emissions of hydraulic excavators, an innovative speed
variable pump controlled and gravitational potential energy recovery integrated boom driving scheme is proposed.
Firstly, the working principle of the proposed system is analyzed. To verify the feasibility and energy saving effectof the proposed system, the test bench of the proposed system are built. The operating and energy efficiency
characteristics of the proposed system are tested and analyzed. The testing results show that, the new system has
good motion characteristics and high positioning accuracy. By using the proposed system, the gravitational potential
energy is recovered as high as 75.3%; compared with the traditional load sensing system, the energy consumption is
reduced by up to 72.4% for lifting and lowering boom each time. In view of good fuel saving and economic
performance, the proposed system can be applied to all kinds of heavy-duty manipulator which is driven by
hydraulic cylinders.
Keywords: hydraulic excavator, three-chamber cylinder, pump controlled system, energy recovery

INTRODUCTION
Hydraulic excavator and other engineering machinery has huge possession around the world, and has
become a major fuel consumption source and pollution emitter [1,2]. With the shortage of fossil fuel, global
warming and strict emission standard implementation, how to reduce the fuel consumption and emissions of thehydraulic excavators has become a research hotspot in this field. The electric hybrid power system compensates
the load fluctuation by the added electric motor/generator, smoothes the engine output torque, and maintains theengine in an high efficient area for a long time, which provides an effective way to solve this problem [3-6].
However, the hydraulic circuit still uses the conventional valve control system to drive multiple actuators. To
complete the same work, compared with the original internal combustion engine power system, the hybrid
power system still needs to provide the same amount of energy [7]. Moreover, the use of the hybrid power systemincreases the investment cost of the hydraulic excavator and the investment cost recovery cycle is long, which
limits the promotion and application of this technology. If the energy efficiency of the hydraulic circuit isimproved, the fuel consumption of the hydraulic excavator can be further reduced and the cost recovery cycle
can be shortened, and the barrier that restricts the application of hybrid power system can be broken.
In recent decades, the proposed load-sensing [8-9], independent metering [10-11] and other technologies canimprove the energy efficiency of the hydraulic system to a certain extent. However, the large throttling loss ofthe valve control system can’t be resolved radically. Taking the hydraulic excavator using load sensing system
as an example, about 45%-50% of the pump output energy is wasted through the throttling effect of the control
valves, and the wasted gravitational potential energy is about 15% of the hydraulic pump output energy [12].
Compared with the valve control system, the pump controlled system doesn’t need control valves and caneliminate throttling loss, and the system has high energy efficiency [13, 14]. Therefore, using pump controlled
systems and efficiently regenerating gravitational potential energy can greatly improve the energy efficiency ofthe hydraulic system.
Recovery and reutilization of the gravitational potential energy is an important way to improve the energy
efficiency of hydraulic excavators. Currently, there are mainly electric and hydraulic two energy recovery
schemes. In the electric energy recovery scheme where the hydraulic circuit is valve control system, thegravitational potential energy is converted to electrical energy by a separately added hydraulic motor-electric
generator, and stored in supercapacitor or battery [15]. In the hybrid hydraulic excavators designed by Komatsu
Ltd [16], Hitachi Ltd [17] and Zhejiang University [18], the gravitational potential energy of the boom is recoveredby the separately added hydraulic motor- electric generator, which has good energy recovery effect. By setting
additional electrical energy storage element, using pump controlled system can realize the unthrottled loss
recovery of gravitational potential energy. Jong et al. studied the characteristics of the excavator’s boom driven

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by a single pump controlled system, in which the potential energy is recovered and stored in supercapacitor andbattery. The energy consumption is reduced by 47.8% [19]. In the hybrid hydraulic excavator developed by the
Kobe Steel Ltd, the boom’s hydraulic cylinder is controlled by a pump controlled system driven by electricmotor. When the boom lowers, the hydraulic pump works in ‘hydraulic motor’ mode and the electric motor
works in ‘electric generator’ mode. In this way, the potential energy is converted into electrical energy andstored in supercapacitor [20]. Chen et al. studied the characteristics of a variable speed pump controlled
excavator’s boom scheme in which the hydraulic accumulator and supercapacitor are used to recovery thegravitational potential energy. The results show that the energy consumption is reduced by 33.1% [21].
In the hydraulic energy recovery schemes, hydraulic accumulator is used as energy storage element. Due tothat the accumulator’ pressure varies with the volume of oil charged, the accumulator has obvious nonlinear
pressure characteristics [22]. To compensate this nonlinearity, during the energy recovery process, the throttling
valve [23], hydraulic transformer [24] or the pump controlled system which recovers the energy by torque couplingshould be adopted. In the hybrid hydraulic excavator proposed by Paolo, a hydraulic accumulator is connectedto the non-rod chamber of the boom’s hydraulic cylinder through a proportional throttle valve. The nonlinearity
characteristics of the hydraulic accumulator is compensated by adjusting the opening of the proportional throttle
valve during the potential energy recovery process. In this scheme, the high pressure oil in the accumulator isintroduced to the inlet of the pilot pump and the recovered potential energy is used to assist the engine to drive
the pilot hydraulic pump [25]. Shen et al. studied the performance of a hydraulic hybrid excavator based oncommon pressure rail. The hydraulic cylinder of each actuator is separately driven by one hydraulic transformer.
Through the hydraulic transformer, the manipulator’s potential energy can be recovered and stored in hydraulicaccumulator with high pressure [26]. However, there is no commercial hydraulic transformer currently, and thesystem is costly and complex [27]. Hippalgaonkar et al. studied the characteristics of a hydraulic hybrid excavator
that each actuator is driven by a single pump controlled system. By adding a hydraulic pump/motor coupledwith engine, the potential and kinetic energy of the excavator can be recovered by means of torque coupling andstored in hydraulic accumulator. Under the same working condition, the fuel consumption is reduced by 50% [28,29]
. A scheme that the hydraulic excavator’s boom is driven by double variable displacement pump controlled
system was proposed by Daniel [30]. In this scheme, a hydraulic accumulator is connected with one hydraulicpump to recover the gravitational potential energy. Ge et al. studied a scheme that the hydraulic cylinder of the
excavator boom is driven by an asymmetric pump [31]. A hydraulic accumulator is connected with one port of
the asymmetric pump to store the high pressure oil of the non-rod chamber of the hydraulic cylinder when theboom falls. The accumulator provides high pressure oil to drive the asymmetric pump when the boom lifts. The
test results show that good energy saving effect is obtained. However, the structure of the asymmetric pump is
complex, and the flow window area ratio is fixed which makes the asymmetric pump only be suitable for thehydraulic cylinder with special area ratio. In the hydraulic energy recovery schemes, the hydraulic-pneumaticbalancing method has the higher energy efficiency due to less energy conversion links [32]. Its basical principle isthat independent energy storage cylinder or chamber is connected with hydraulic accumulator to recover andreutilize the gravitational potential energy through the way of balancing the self-gravity of working device.
Liang firstly theroretically studied the characteristics of a loglift loader’s boom driven by two energy storage
cylinders and one working cylinder [33]. In reference [34], a scheme that the hydraulic excavator boom is driven
by one energy storage cylinder and two working cylinders is presently, and the energy consumption is reducedby 26.2%. Currently, the scheme adopting the energy storage cylinder has been applied to the boom of material
handling machines. To apply this method in the compact installation space manipulator driven by single
hydraulic cylinder, a hydraulic-pneumatic balancing scheme using three-chamber cylinder having an energy
storage chamber is proposed in reference [35, 36], and good energy saving effect is obtained [37].
Through the above analysis, we can see that, in the existing hybrid hydraulic excavators, a separate hydraulicmotor-generator unit is used to recover the gravitational potential energy, which increases the systemcomplexity, cost, and the cost recovery cycle, limiting its popularization and application. Moreover, thegravitational potential energy is recovered through conversion links of the hydraulic motor, the generator, theDC-DC converter and supercapacitor or battery, which lowing the energy recovery efficiency. During theenergy reutilization process, there is a large throttling loss because of the original valve controlled system,
decreasing the overall energy saving effect. By using the pump controlled system, the gravitational potential
energy can be conveniently converted into electrical energy and stored in supercapacitor, or be converted into
hydraulic energy through torque coupling and stored in the accumulator. However, there are many energy
conversion links, which reduces the recycling efficiency of the gravitational potential energy. Moreover, thevariable speed or displacement pump controlled system uses two pumps, the structure of the systems is complex
and the cost is high. For the large area difference of the differential hydraulic cylinder, in the single pumpcontrolled system, a filling system with large flow rate is needed which causes large energy loss. In the existingenergy recovery schemes, the hydraulic-pneumatic balancing method has a high gravitational potential energy
recovery and reutilization rate, but the main hydraulic circuit is still valve control system, resulting in a large
throttling loss [34].

2

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Based on the feature that hybrid electric excavator has electric energy storage components such as super
capacitors and batteries, to address above issues, an innovative speed variable pump controlled and gravitational
potential energy recovery integrated boom driving scheme is proposed in this paper. The core element of thesystem is a three-chamber cylinder. Two chambers of the three-chamber cylinder are directly connected to thehydraulic pump to drive the excavator boom without throttling loss. The other chamber (called energy storage
chamber) of the three-chamber cylinder is connected with a hydraulic accumulator. When the boom is lowered,
the high-pressure oil of the energy storage chamber is charged into the accumulator under the gravity of theworking device to efficiently recover the gravitational potential energy. When the boom is lifted, theaccumulator releases high-pressure oil to the energy storage chamber to drive the boom, which realizes theefficiently reutilization of the recovered energy. The proposed system also has the following advantages. For
small area difference between the non-rod chamber and the rod chamber, an oil filling system with small flowrate is needed which reduces the system energy loss. Because the self-gravity of the working device is balancedby the energy storage chamber, the hydraulic pump only needs to drive the load and compensate the inertia force,
which greatly reduces the installed power compared with the above mentioned pump controlled system.
The paper is organized as follows. In section two, the working principle of the proposed system is illustrated andthe velocity feedforward and displacement feedback (VFDB) compound control strategy is designed. In section
three, the energy efficiency mathematic models of the proposed system and the valve controlled single rod
cylinder system are built. In section four, the operation performance and energy saving effect of the proposedsystem is tested and analyzed. In section five, the conclusiton is given.

WORKING AND CONTROL PRINCIPLE
Working principle
The working principle of the proposed system for excavator boom is shown in Figure1. The system consistsof a controller, a servo-motor, a servo driver, a hydraulic pump, a filling system, a hydraulic control directional
valve, a hydraulic accumulator, and a three-chamber cylinder and so on. As shown in Figure1, by adding a fixedpiston and manufacturing the piston rod with hollow, the single-rod hydraulic cylinder can be reformed as athree-chamber cylinder. The three-chamber cylinder has non-rod chamber A, rod chamber B and energy storage
chamber C. The chamber A and chamber B are connected with the hydraulic pump, respectively. The velocity
and displacement of the boom is controlled by adjusting the rotation speed and direction of the servo-motor,
respectively. The energy storage chamber is connected with the hydraulic accumulator. The oil in the energy
storage chamber is pressed into the hydraulic accumulator under the action of the working device’s self-weight
when the boom lowers, and the gravitational potential energy is recovered without throttling loss. When theboom lifts, because of the high oil pressure in the hydraulic accumulator, the energy storage chamber outputs
force to assist to drive the boom. By this way, the recovered potential energy can be directly reutilized.

Arm
7
Boom

BucketA

B
B

7

Pistonrod
FixedpistonCylinder
body

1

5b

4

3
2

M

6
M

u

B

Servo
driver5aCC

vset, ssetx
A

Controller

u

C

p
u

p

1. Servo-motor 2. Hydraulic pump 3. Oli filling system 4. Hydraulically operated directional control valve
5a, 5b Relieve valve 6. Hydraulic accumulator 7. Three-chamber cylinder
Figure 1 Principle of the proposed system for excavator boom

In the previous study [37], the parameters of the hydraulic accumulator and the three-chamber cylinder has
been designed. The capacity of the hydraulic accumulator is selected as 20 L and initial gas pressure is set as 5.5
MPa. Table 1 shows the chamber areas of the hydraulic cylinders. It’s known from Table 1 that the area

3

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difference between the non-rod chamber and rod chamber of the original working cylinder is about 4900 mm2
and that of the three-chamber cylinder is about 1500 mm2. Compared with the original working cylinder, thearea difference of the three-chamber cylinder is reduced by 2.26 times, which can greatly reduce the installed
power of the filling system. In this paper, the flow rate of the filling system is about 15 L/min and the pressure is
2 MPa. Calculation shows the power of the oil filling system is 0.5 kW. If the single-rod hydraulic cylinder isused, the power of the oil filling system would be 1.6 kW.
Table 1. Areas of hydraulic cylinders

Three-chamber cylinder
Single-rod hydraulic cylinder

Chamber A
1963.5 mm27850 mm2

Chamber B
3141.6 mm23846.7 mm2

Chamber C
7539.8 mm2

Control strategy
The proposed system has typical four-quadrant operation characteristics as shown in Figure 2 When thesystem works in the first quadrant or the fourth quadrant, the pressure of the chamber A is higher than that ofchamber B, which makes the chamber A as the control chamber.
Ⅰ

Ⅲ

Ⅱ

qB

Ⅳ

qB

qB

qB
qfill

qC

n

qC

n

qC

qfill

qA

qA

n

n
qA

qC

qA

pA  pB v > 0

pB  pA v > 0

pB  pA v < 0

pA  pB v < 0

qp =AA·v

qp = AB·v

qp = AB·v

qp = AA·v

Figure 2 Four-quadrant principle of the proposed system

The output flow of the hydraulic pump is calculated according to the area of the chamber A and the velocityof the three-chamber cylinder.
qP = AA v(1)
where v is the velocity of three-chamber cylinder.
When the system works in the second quadrant or the third quadrant, the pressure of the chamber A is lower
than that of chamber B, which makes the chamber B as the control chamber. The output flow of the hydraulicpump is calculated according to the area of the chamber B and the velocity of the three-chamber cylinder.
qP = AB v(2)
The three-chamber cylinder velocity control signal us is set as control signal of the velocity feedforward
control unit, the rotation speed of the servomotor is shown as follows.
n = k s us
(3)
where n is the servomotor rotation speed, ks is the gain.
Ignore the leakage of the hydraulic pump, the flow rate of the hydraulic pump is:(4)
qp = nDp
where Dp is the displacement of the hydraulic pump.
According to Eqs. (1), (2), (3) and (4), the relation between the v and us is shown as follows.
 AA v(pA  pB )
D k
 p s
us = 
(5)
 AB v(pB <pA ) Dp ksIn the velocity feedforward control unit, the leakage of the hydraulic pump is not considered. There will be alarge deviation of the expected velocity and the actual velocity. Therefore, a displacement feedback closed
control unit is added. The integration of the velocity control signal us is set as the expected displacement signal
sin in the displacement feedback unit. In this way, the comparative value of the expected displacement signal andthe actual displacement signal is small. The feedforward control unit plays a major role and the displacement
feedback control unit plays an auxiliary regulation role. The three-chamber cylinder can be operated according
to set velocity and displacement.

4

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The difference between the actual displacement and the integrated displacement is:s =  vdt − sact(6)

where sact is the actual displacement.
The PI controller is used in the displacement feedback control unit, the rotation speed compensation signal of
the servomotor is:kp
uc = kp s +  sdt
(7)
Ti
where kp and Ti are respectively the gain and integration time constant of the PI controller.
Figure 3 shows the principle of the VFDB control strategy. According to the detected pressures, the velocity
feedforward control unit determines which chamber is the control chamber, and then outputs the control signal
us. Meanwhile, the displacement feedback control unit outputs the control signal uc according to thedisplacements difference. The controller outputs the sum of us and uc to control the servomotor rotation speed,
and then the pump outputs flow according to the demand.
vset
us

v
0

0

0

0

0

pA

0

sset

＋

Curve generator

1
s
＋

＋ uc

1
Ti s

kp
sin

AB
Dp ks

no

pA

sact

vact

AA
Dp ks

＋

kp

F

yes

＋

－

Boom

pB?

n

pB

Threechamber
CylinderpHydrualic
Pump

qp

n
T

Servo
Motor

Figure 3 Principle of velocity feedforward and displacement feedback control strategy

Where vset is the setting maximum velocity, sset is the setting object displacement, T is the servomotor torque,p is the pump pressure, F is the three-chamber cylinder force, vact is the actual velocity.

ENERGY EFFICIENCY MATHEMATICAL MODEL
To describe the energy saving effect of the proposed system, the traditional load-sensing system is adopted
as a comparison. The pressure difference compensator is used to keep the pressure difference between the inlet
and outlet of the directional valve constant, and the output pressure of the hydraulic pump is always about 2-3
MPa higher than the load pressure. Figure 4 shows the working principle of traditional load-sensing system.
x,v
u

Boom

p

dSPACE 1103

u

B
p

u

A

LS valveFlowsensorJoystick

p

Pump

u

E Engine

Figure 4 Working and test principle of the traditional load-sensing system

For the traditional load sensing system, the hydraulic pump outputs high pressure oil into the non-rod
chamber A through the load sensing valve (LS valve) to lift the boom, and the oil in the rod chamber B flows to
the tank. The output energy of the hydraulic pump is：

Eup =  pp AA vdt
t

0

(8)

where AA is area of the non-rod chamber A, pp is pressure of the hydraulic pump, t is the boom lifting time;5

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The energy loss of the system is the throttling loss of the LS valve, which is shown as follows.

Euploss =  v( AA ( pP − pA ) + AB pB )dt
t

0

(9)

where AB and pB are the area and pressures of the rod chamber B, respectively; pA is the pressure of non-rod
chamber pressure.
Similarly, when the boom lowers, the output energy of the pump and the throttling loss of the LS valve areshown respectively:

Edown =  pP AB vdt(10)

Edownloss =  ( AB ( pP − pB ) + AA pA )vdt

(11)

t0

0

t0

0

where t0 is the boom lowering time.
The energy consumed by the load sensing system to drive the boom for one cycle is shown as follows.
E = Eup + Edown(12)

Based on the energy conservation law, the recoverable potential energy of the working device is:

Ep =  v( AA pA − AB pB )dt
t0

0

(13)

For the proposed system, to lift the boom, the pump only outputs energy in the first quadrant. The energy
consumption is: =  AA v( pA − pB )dt
Eup
t1

0

(14)

where t1 is the first quadrant time.
During the boom lifting process, the energy loss is the energy discharged to the filling system by the rod
chamber.
t1

t − t100Euploss=  vpB ( AB − AA )dt + (vAB ( pB − pA ) − vpA ( AB − AA ))dt

(15)

Similar to the lifting process, the consumed energy of the proposed system to lower the boom is shown as
follows.

 =  [( AB pB − AA pA )v+pA (AB -AA )v]dt
Edown
t2

0

(16)

where t2 is the third quadrant time.
The energy loss is:
Edownloss=

t0 − t2v( AA ( pA − pB ) + pB ( AB − AA ))dt

(17)

The total pump controlled system energy consumption is shown as follows.
 + EdownE  = Eup(18)0

During the operating cycle of the boom, the thermodynamic equation of the hydraulic accumulator is shownas follows.
(19)
pxVxk =p0V0k =const
where px is and Vx are the gas pressure and volume of the hydraulic accumulator at any time, respectively; p0
and V0 are the initial gas pressure and volume of the hydraulic accumulator, respectively; k=1.4 which represents
gas index.
When the boom lowers, the energy stored by the hydraulic accumulator is calculated by Eq. (20).
t0p2V2 k AC v
 =
Eacc
dt
(20)
0 (V − A vt ) k
2Cwhere p2 and V2 are the gas pressure and volume of the hydraulic accumulator when the boom begins tolower, respectively.
When the boom lifts, the energy supplied by the hydraulic accumulator is calculated by Eq. (21).
tp1V1k AC v
Eacc = dt(21)
0 (V + A vt ) k
1
C
where p1 and V1 are the gas pressure and volume of the hydraulic accumulator when the boom begins to lift,
respectively.
Besides the potential energy, the energy stored by the hydraulic accumulator also includes the output energyof the hydraulic pump. The recovered potential energy is calculated according to the following equation.

Er = Eacc −  ( AB pB v − AA pA v)dt
t0

0

(22)

The potential energy recovery rate is:6

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r =

Er
Ep

By using the proposed system, the reduced energy consumption is:E  +E s = 1- up down
Eup +Edown

(23)

(24)

EXPERIMENT RESEARCH
Testing system
The testing platform of a 6-ton hydraulic excavator’s boom controlled by the traditional load sensing system
was built firstly, and the testing principle is shown in Figure 4. To analyze the feasibility, operation performanceand energy characteristics of the proposed system, the testing rig of the proposed system is built, and the testing
system is shown in Figure 5. In the testing systems, the Parker flowmeter is used to collect the signals of thepressure and flowrate of the hydraulic pump. The Atos pressure sensor is used to detect the pressures of thehydraulic cylinders. The MTS magneto-strictive displacement sensor is used to detect the displacement andvelocity of the hydraulic cylinder. A braking resistor is used to balance the generated electrical energy of theservomotor. The dSPACE 1103 is used as controller and data acquisition.
AC
380V

vset
sset

Controller

F
Boom

u
xv

up up

Driverup

B
M
3~~
ServomotorC(a) Testing principle(b) Testing picture
Figure 5 Testing principle of the proposed system

Operation characteristics
The influence of the hydraulic accumulator pre-charge pressure on the four-quadrant characteristic was
firstly analyzed. Under the same working condition, Figure 6a and Figure 6b show the servomotor bus voltage
and the system energy consumption under different hydraulic accumulator pre-charge pressures, respectively.
The bus voltage is about 540 V when the servomotor works in motor condition. If the servomotor works undergeneration condition, the generated electrical energy will make the bus voltage higher than 540 V. When thehydraulic accumulator pre-charge pressure is 10 MPa, the output force of the energy storage chamber exceeds
the gravity of the hydraulic excavator working device in the ab section. The energy supllied by the hydraulicaccumulator is larger than the required energy to lift the boom. Thus, a part of the supllied energy of thehydraulic accumulator converts into electrical energy through the servomotor, and the bus voltage rises to nearly
600 V. As shown in Figure 6b, the energy consumption of the hydraulic system is the highest. When the initialpressure is 7 MPa, the output force of the energy storage chamber cannot balance the self-gravity of the working
device in the cd section. The hydraulic accumulator cannot completely recover the potential energy of theworking device and a part of the potential energy is converted into electrical energy and wasted through thebraking resistance, and the bus voltage rises to nearly 590 V. The energy consumption of the hydraulic system isabout 0.94 of the system under 10 MPa hydraulic accumulator pre-charge pressure.

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600

200

c

Voltage U/V

580

100

Setting
velocity560

0

10 MPa
7 MPa

540
8 MPa

520

0

3

6
Time t/s

9

1

1

d

0.96

E/Emaxb

Velocity v/(mm/s)

a

0.94

0.92

-100

0.88

-200

0.84

12

0.91

1
7 MPa

2
8 MPa

3
10 MPa(a) Bus voltage of servomotor(b) Energy consumption of hydraulic system
Figure 6 Characteristics of the proposed system with different accumulator pre-charge pressures

It should be noted that the bus voltage of the servomotor is always about 540 V when the pre-charge pressureis 8 MPa. It means that the servomotor is always working in motor condition and the impact of the negative loadon the system can be avoid, which can improve the control performance of the proposed system. It also meansthat the hydraulic accumulator can recover the potential energy as much as possible and the supplied energy ofthe hydraulic accumulator isn’t wasted. It can be seen in Figure6b that the energy consumption of the hydraulicsystem is the lowest. To keep the consistency, the following test results of the proposed system are all obtained
under 8 MPa pre-charge pressure of the hydraulic accumulator.
Figure7a shows the tested velocity and displacement of the three-chamber cylinder adopting VFDB closed
loop control strategy. During the testing process, the set maximum velocity vset is 80 mm/s and the set object
displacement sset is 200mm. By using the VFDB closed loop control strategy, the velocity setting value changes
along with the displacement difference for the compensation of displacement feedback control unit. The
positioning error is 1 mm. Figure 7b shows the pressures and displacement of the three-chamber cylinder.
During the boom operation process, the actual displacement of the three-chamber cylinder can meet therequirement of the expected setting displacement, indicating the system has good repeatability. Because of thelarger area of the energy storage chamber, when the piston rod extends, the small pressure decreasing in theenergy storage chamber makes the pressure of the non-rod chamber increase greatly. Meanwhile, the rod
chamber discharges oil into the filling system and the rod chamber pressure equals to the filling system pressure
which is about 2 MPa. To lower the boom, the rod chamber pressure keeps high pressure whose mean value isabout 6.3 MPa. Because the oil filling system charges oil into the non-rod chamber, the pressure of the non-rod
chamber is 2 MPa.

600

0

500

Actualdisplacement

450

Setting
displacement

400

-100

350

0

3

6

9
Time t/s

1215

Actual Displacement

16

600

12

pA

pC
400

8

pB
200

4
0

Displacement s/mm

1 mm

550

-50

800
Setting Displacement

Pressure p/MPa

Actualvelocity

50

20

650

Setting velocity

Displacement s/mm

Velocity v/(mm/s)

100

0

0

15

Time t/s

3045(a) Velocity and displacement (b) Displacement and pressures
Figure 7 Velocity, displacement and pressures of the proposed system

Energy saving characteristics
Figure 8 shows the pressures and energy losses of the different systems when the boom moves 330 mm. The
pressures of the traditional load sensing system and the proposed system are respectively shown in Figure 8a
and Figure 8b. As shown in Figure 8a, the pressure of the pump outlet port is about 3 MPa higher than that of
the hydraulic cylinder in the traditional load sensing system. There are large pressure fluctuation during thelifting process and the non-rod chamber pressure is about 9.5 MPa. To provide enough force to stability drive
the boom lowering, the non-rod chamber keeps 7 MPa pressure. The hydraulic accumulator pressure and8

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capacity in the proposed system are selected according to the above pressure values of the single rod cylinder.
As shown in Figure 8b, the pressure curves of the proposed system are the same as those shown in Figure 7. For
the gas volume of the hydraulic accumulator decreases when the boom lowers, the pressure of the hydraulicaccumulator increases which causes the oil pressure rise in the rod chamber of the three-chamber cylinder.
As shown in Figure 8b, the highest pressure in non-rod chamber is about 16.5 MPa under the bucket without
load condition. In general, the hydraulic system’s maximum pressure of the hydraulic excavator is 28 MPa. Dueto the gravity of the load is far lighter than that of the working device, it’s known from the previous study [38]that the highest pressure of the non-rod chamber is about 21 MPa under the bucket with full-load condition,
which means the pressure level can meet the requirements. Meanwhile, the designed parameters and selectedmaterial of the three-chamber cylinder also can meet the structural strength requirements.
700

Displacement

Weight-balancechamber pressureRodchamberpressure

15

600

10

500
Rodless chamberpressure

5

400

0

Displacement s/mm

Pressure p/MPa

20

300
0

3

6

9

12

Time t/s(a) Traditional load-sensing system(b) Proposed system
Figure 8 Pressures of the different systems

Figure 9a and Figure 9b show the energy losses of the traditional load-sensing system and proposed system.
As shown in Figure 8a, there is a large pressure loss between the inlet port and outlet port of the control valve
due to the throttling effect. It’s noted from Figure 9a that a large amount of energy is wasted in the traditionalload sensing system. When the boom lifts, the energy loss is mainly the output energy of the hydraulic pump,which is about 12.6 kJ. During the lowering process, the energy loss mainly consists of the output energy of thehydraulic pump and the wasted potential energy, between which the energy loss of the hydraulic pump is about
4.5 kJ. The wasted potential energy of the working device is about 16.5 kJ which accounts for a large parts ofthe energy loss. Therefore, if the gravitational potential energy can be recovered and reutilized, the energy
consumption can be significantly reduced. In the proposed system, the energy loss mainly consists of thedischarging energy of the three-chamber cylinder as shown in Figure 9b. Because the discharging flow rate issmall and the pressure is low, the energy loss in the proposed system is very small. The energy loss is about 1.2
kJ during the lifting process and 0.1 kJ during the lowering process.
16

0.4
Pump output power loss

Recoverable potential
energy power

8

0.3

Power P/kWPower P/kW

12

Pump output
power loss

0.2
0.1

4

0

0
0

2

4

6
Time t/s

8

10

12

0

3

6
Time t/s

912(a) Traditional load-sensing system(b) Proposed system
Figure 9 Energy losses of the different systems

Figure 10a and 10b show the powers and energy consumptions of the traditional load-sensing system andproposed system, respectively. The peak power of the traditional load-sensing system is about 13.5 kW for
lifting the boom. Conversely, the peak power of the proposed system is about 3.3 kW for lowering the boom asshown in Figure 10b, which means the peak power is reduced by 76.1% compared with the traditional loadsensing system.
When the boom lifts, the output power of the proposed system is obvious less than that of the traditionalload-sensing system as shown in Figure10. The hydraulic pump in the traditional load-sensing system outputs

9

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about 11.5 kW mean power and 38.2 kJ energy to lift the boom. During the lifting process, the maximum pump
output power of the proposed system is about 2.2 kW and the supplied power of the hydraulic accumulator isabout 5.2 kW. The hydraulic pump outputs 6.7 kJ energy and the hydraulic accumulator supports 20.2 kJ energy
to lift the boom. The energy supplied by the hydraulic accumulator accounts for 75.4% of the required energy tolift the boom. To lower the boom, the rod chamber pressure of the three-chamber cylinder is high as shown in
Figure 8b. The rod chamber outputs large force to fill the oil of the energy storage chamber into the hydraulicaccumulator. It is seen from the Figure 10 that the hydraulic power of the proposed system is larger than that of
traditional load-sensing system. For the lifting and lowering distances are the same, the supplied energy andstored energy of the hydraulic accumulator are basically identical.

Peak power
13.5 kW
15

10

30

5

15
Pump output
power

0
0

3

6
Time t/s

09Power P/kW

Recoverable
potential energy

Accmulator supply power

445Energy E/kJ

Power P/kW

860
Pump output
energyAccumulator supply
energy25

20

Peak power
3.3 kW15
0

Pump
output
power

-4

10

5
Pump output
energy

-8
0

12

2

Energy E/kJ

20

4

6
Time t/s

0
8

10

12

(a) Traditional load-sensing system(b) Proposed system
Figure10 Power and energy consumption of different systems

Table 2 shows the testing results of the traditional load-sensing system and the proposed system according to
the energy model in section 3 and the testing data.
Table 2 Testing results

Energy loss (kJ)
Recoverable potential energy (kJ)
Hydraulic accumulator recovered energy (kJ)
Recovered potential energy (kJ)
System peak power (kW)
System total energy consumption (kJ)

Traditional load-sensing system
17.1
16.5
0
0
13.8
55.5

Proposed system
1.3
16.5
20.85
12.35
3.3
15.3

It’s known from the Table2 that about 16.5 kJ gravitational potential energy can be recovered when theboom lowers. In the proposed system, the output energy of the hydraulic pump is about 8.5 kJ to lower the boom
and the energy stored by the hydraulic accumulator is about 20.85 kJ. According to Eq. (22), about 12.35 kJ
gravitational potential energy is recovered by the hydraulic accumulator, which is about 75.3% of therecoverable gravitational potential energy. According to the previous analysis, the gravitational potential energy
can be almost completely recovered under 8 MPa hydraulic accumulator pre-charge pressure. However, the test
results are not very satisfactory. The main reason is that the friction forces in the three-chamber cylinder and thehinged joints are larger. To drive the boom, the traditional load sensing system consumes about 55.5 kJ energy,
and the proposed system consumes about 15.3 kJ energy. Compared with the traditional load-sensing system,
the energy consumption is reduced by 72.4%.
In the previous research work, the characteristics of load-sensing system with three-chamber cylinder driving
excavator boom were studied [36]. During the period of the boom lifting and lowering 330 mm, the energy
characteristics of the different systems are shown in Figure 11. As can be seen, the traditional load-sensing
system has the largest energy loss and energy consumption, and the highest peak power. The use of a threechamber cylinder instead of a single-rod hydraulic cylinder in the load-sensing system significantly reduces theenergy consumption, energy loss and peak power of the boom system. It’s noted that the proposed systemfurther reduces the boom system energy consumption by 27.5%, energy loss by 69%, and peak power by 28%
compared to the load-sensing system with three-chamber cylinder. Since the accumulator volume and the precharge pressure are the same, the gravitational potential energy recovery efficiencies of the proposed system andthe load-sensing system with three-chamber cylinder are basically the same.

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Energy or Power E or P/kJ or kW

60

55.5
C

B

A
45

30
21.1
15.3

17.1
12.4 12.35

15
4.2

1.3

13.8
4.6 3.3

00Energy loss

RecoveredenergyEnergyconsumption

Peak power

A-Traditional load sensing system, B-Load sensing system with three-chamber cylinder, C-Proposed system
Figure 11 Energy characteristics of different systems

Economic and environmental benefits
Although the energy saving effect of the proposed system cannot directly reflect the reduction of the fuelconsumption of the internal combustion engine, the fuel saving effect of the proposed system can be evaluatedaccording to the fuel consumption rate of the engine.
Assuming in the 90°excavation cycle, the distance between the excavation point and the unloading point is
330 mm, meaning that the boom falls and lifts 330 mm, respectively. According to the above calculation, theoutput energy of the hydraulic pumps is reduced by 33 kJ for one excavation cycle. Considering the efficiencyof the hydraulic pump is about 90%, during one 90°excavation cycle, the reduced energy converted onto theengine axis is 36.7 kJ which represents the reduced output energy of the engine.
During the actual operation process, the time for operating a 6-ton hydraulic excavator to complete one 90°
excavation cycle is 15 s. Thus, the hydraulic excavator can complete 240 cycles per hour and 8808 kJ energy
can be saved per hour. According to that the average fuel consumption rate of the engine is 0.35 L/kWh, the fuelcan be saved by 0.86 L per hour. Considering the current fuel price about 7.1 RMB/L in China, it can be
estimated that the equivalent fuel cost should be saved about 6.1 RMB per hour by using the proposed system.
The total cost of the prototype of the proposed system is estimated about 23000-25000 RMB, among which
the cost of the permanent magnetic synchronous machine and its driver is 14000 RMB, the three-chambercylinder is 5000 RMB, the hydraulic accumulator is 1200 RMB and the other components accounts for 28004800 RMB. If the mass production of the proposed system is realized, the cost can be further decreased.
Assuming that the hydraulic excavator is operated for 10 hours per day and 300 days per year, the operation
time is 3000 hours per year. By using the proposed system, the running cost of one 6-ton hydraulic excavator
can be reduced by 18300 RMB. Thus, after one and a half years, the upfront investment of the proposed system
could be recovered by the revenue of fuel saving, and one 6-ton hybrid hydraulic excavator using the proposedsystem can bring about 18300 RMB per year for customer.
According to the fuel saving per hour and operation time per year, by using the proposed system, the fuel canbe saved by 2569 L per year. The carbon dioxide emission of diesel oil is 2.63 kg/L, and one 6-ton hydraulicexcavator can reduce the greenhouse gas emissions by 6756.5 kg per year.
Considering the rapid recovery of the upfront investment, the great potential of energy saving and emission
reduction, the innovative scheme proposed in this paper can be applied to all heavy-duty manipulators driven by
hydraulic cylinder. It not only can be used to manufacture new hydraulic excavators, but also can be used toreform the existing hydraulic excavators through designing the proposed prototype as retrofit kit.
Considering the rapid recovery of the upfront investment, the great potential of energy saving and emission
reduction, the innovative scheme proposed in this paper can be applied to all heavy-duty manipulators driven by
hydraulic cylinder. It not only can be used to manufacture new hydraulic excavators, but also can be used toreform the existing hydraulic excavators through designing the proposed prototype as retrofit kit.

CONCLUSIONSThe novel scheme of speed variable pump controlled three-chamber cylinder system can realize theintegration of the driving and potential energy recovery and reutilization of the hydraulic excavator, andcompletely eliminate the throttling loss. The scheme has the advantages of shorter energy transfer chain, lessenergy conversion links, and high gravitational potential energy recovery and reutilization rate. The novel

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scheme is not only suitable for the speed variable pump controlled system, but also for the displacement variable
pump controlled system.
The testing results show that the pre-charge pressure of the hydraulic accumulator has great influence on thefour-quadrant characteristics of the proposed system. For the 6-ton hydraulic excavator, when the pre-chargepressure of the hydraulic accumulator is 8 MPa, the servomotor always works in resistance load condition
whether the boom lifts or lowers, and the energy efficiency of the system is the highest under above mentioned
condition.
The proposed system has good operating characteristics. The VFDB control strategy is used to improve thetrajectory tracking precision of the hydraulic excavator boom, and the positioning error is reduced to about 1mm.
By using the proposed system, the gravitational potential energy recovered by the proposed system can reachto 75.3% of the recoverable energy. To lift the boom, 75.4% of the required energy is supplied by the hydraulicaccumulator. Compared with the traditional load-sensing system, the peak power is reduced by 76.1% and theenergy saving ratio of the proposed system is 72.4% under the same working condition. This new principle not
only reduces the energy used to do useless work, but also greatly reduces the installed power of the main drive
circuit.

ACKNOWLEDGMENTS
The authors would like to give their acknowledgment to National Natural Science Foundation of China (No.
51775363, 52105066 and U1910211) for the financial support on the project.

REFERENCES
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cycles as a priori. Energy 2017; 133: 929-940.
6. Liukkonen M, Lajunen A, Suomela J. Feasibility study of fuel cell-hybrid powertrains in non-road mobile machineries.
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7. Lin T, Wang Q, Hu B, et al. Development of hybrid powered hydraulic construction machinery. Automation inconstruction 2010; 19(1): 11-19.
8. Xu B, Cheng M, Yang H, et al. A hybrid displacement pressure control scheme for an electro-hydraulic flow matchingsystem. IEEE/ASME Transactions on Mechatronics 2015; 20(6): 2771-2782.
9. Axin M, Eriksson B, Krus P. A flexible working hydraulic system for mobile machines. International Journal of Fluid
Power 2016; 17(2): 79-89.
10. Huang W, Quan L, Huang J, et al. Excavator Swing System Controlled with Separate Meter-in and Meter-out Method.
Journal of Mechanical Engineering 2016; 52(20): 159-167.
11. Ding R, Xu B, Zhang J, et al. Self-tuning pressure-feedback control by pole placement for vibration reduction of
excavator with independent metering fluid power system. Mechanical Systems and Signal Processing 2017; 92: 86-106.
12. Zimmerman J, Pelosi M, Williamson C, and Ivantysynova, M. Energy Consumption of an LS Excavator Hydraulic
System. In: ASME 2007 International Mechanical Engineering Congress and Exposition, Washington, USA, 11-15
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19. Yoon J I, Truong D Q, Ahn K K. A generation step for an electric excavator with a control strategy and verifications ofenergy consumption. International Journal of Precision Engineering & Manufacturing 2013; 14(5): 755-766.
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21. Chen M, Zhao D. The gravitational potential energy regeneration system with closed-circuit of boom of hydraulicexcavator. Mechanical Systems & Signal Processing 2017; 82:178-192.
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34. Hao Y, Quan L, Cheng H, et al. Potential energy directly conversion and utilization methods used for heavy duty liftingmachinery 2018; 155(2018): 242-251.
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37. Xia L, Quan L, Ge L, et al. Energy efficiency analysis of integrated drive and energy recuperation system for hydraulicexcavator boom. Energy Conversion and Management 2018; 156: 680–687.
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hydraulic excavator boom. Journal of Mechanical Engineering 2017; 53(20): 176-173.

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Experimental Validation of Improvement of the Overall Efficiency
for Electro-hydraulic Drive System using Efficiency Maps
Ha Tham PHAN* and Yasukazu SATO***

Graduate School of Engineering Science, Yokohama National University(E-mail: ha-phan-dz@ynu.jp)**
Faculty of Engineering, Yokohama National University(E-mail: sato-yasukazu-zm@ynu.ac.jp)
79-5 Tokiwadai, Hodogaya, Yokohama, Kanagawa, Japan

Abstract. In industrial applications, electro-hydraulic drive systems (EHDS) included electric motor driven
hydraulic pump are preferred to mechanical and electrical drive system for their advantages. Recently, a crucial
research topic in EHDS is to improve the overall energy efficiency. To do so, the valveless EHDS (EHDS without
control valve) is proposed by researchers and the advantages of valveless EHDS regarding its energy efficiency isvalidated. In previous published research, the novel control method of EHDS proposed by authors is to use overall
efficiency maps combined from motor efficiency map and pump efficiency map for controlling motor speed andhydraulic pump displacement simultaneously to force EHDS operating at high-efficient points. The simulation
results shows that the efficiency of two-degree control system (VS-VP: Variable Speed Motor-Variable
Displacement Pump) has improved more than one-degree control system. This paper presents the validation of thismethod by experiment in actual machines.
Keywords: Electro-hydraulic drive system, Servomotor, Hydraulic pump, Efficiency map

INTRODUCTION
The valveless electro-hydraulic drive system typically includes an electric motor and hydraulic pump to supply
hydraulic power to an actuator as shown in Figure 1. The flow rate supplied to actuator will be regulated bychanging only the speed of electric motor (VS-FP: variable speed-fixed displacement), controlling only thedisplacement of hydraulic pump (FS-VP: fixed speed-variable displacement), or changing motor speed andpump displacement simultaneously (VS-VP). Therefore, the efficiency of EHDS depends on the efficiency ofelectric motor and hydraulic pump.

Figure 1. Valveless electro-hydraulic drive system

The formulas to determine the efficiencies of electric motor and hydraulic piston pump have been presented in
the previous study as shown in (1) and (2). In case of servo motor, energy efficiency is determined bymechanical power at motor shaft and electrical power supplied to servo amplifier [1].

m 

Pout
Tm nm


Pin Tm nm  PCu  PFe  PC

Tm nm
2

T 
T
Tm nm  3RM  m   K f 1nm  K f 2 nm2  K C m
KT
 KT 

(1)

where, Tm, nm are torque and speed at motor shaft; PCu, PFe, PC are copper losses, iron losses and current
losses at electric motor. RM is motor winding resistance; KT, Kf1, Kf2, and KC are experimental coefficients.
The total efficiency p of variable dipslacement pump such as hydraulic piston pump is determined fromvolumetric efficiency pv and hydraulic-mechanical efficiency pm as expressed in Equation (2) [2,3].






C s p  1 p   pv pm  1 


2


n



T
2

2
1
 p
 
c
n p  1   Cv1 2  Cv 2  
 Cf 
 

60

 


p
pD
60


max
 


(2)

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Motor torque Tm (Nm)Pressure differential p (MPa)where Cs, Cv1, Cv2, Cf and =D/Dmax are non-dimensional quantities. , p, Ts, np and Dmax are dimensional
quantities with the unit of Pa.s, Pa, Nm, rpm and m3/rev, respectively.
From theoretical formulas (1) and (2), the efficiency of servo motor and hydraulic pump can be determined andused to create the efficiency maps. For variable displacement pump, the number of efficiency maps will dependon the displacement ratio . The overall efficiency maps of EHDS are combined from servo motor efficiency
map and hydraulic piston maps. The theoretical efficiency maps of servo motor and hydraulic piston pump in
previous study are shown in Figure 2 and Figure 3, respectively [2].

Motor speed nm (rpm)Figure 2. Efficiency map of servo motor

Flow rate Q (L/min)Figure 3. Efficiency map of hydraulic piston pump with  = 1

In previous research [2, 4], the novel control method of EHDS proposed by authors is to use overall efficiency
maps combined from motor efficiency map and pump efficiency map for controlling motor speed and hydraulicpump displacement simultaneously to force EHDS operating at high-efficient points. The numerical simulation
was conducted for 3 types of EHDS (VS-FP, FS-VP and VS-VP) at different operating points (pressure
differential p, flow rate Q). The simulation results show that the efficiency of two-degree control system (VSVP) was improved significantly compared with conventional one-degree control systems (VS-FP, FS-VP). The
advantage of VS-VP system was achieved not only separated working points but also entire operating range. In
this study, an experimental validation of two-degree control system is conducted in actual machines.

TWO-DEGREE-OF-FREEDOM CONTROL OF EHDSIn EHDS, the flow rate Q supplied to actuator is determined by the following equation:
Q  nm Dmax pv(3)

From formula (3), the pump speed nm and displacement ratio  could be controlled to reach command flow rate
Q. However, from formulas (1) and (2), for a given working point (pressure p, flow rate Q) the efficiency ofelectric motor and hydraulic pump are depended on pump speed nm and displacement ratio . In other word, theefficiency of EHDS is a function of pump speed nm and displacement ratio .

Figure 4. Two-degree-of freedom control system

2

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To control EHDS with high-efficient operation for required pressurep and flow rate Q, a two-degree-offreedom control strategy was proposed in authors’ previous research as shown in Figure 4. The target of thissystem is to optimize overall efficiency of EHDS by using two degrees of freedom (motor speed nm and pumpdisplacement ratio ) for specified given working conditions (p, Q) based on the data of the overall efficiency
maps. In other words, for a given value of p, Q the controller will determine the value of efficiency in every
single overall efficiency map ( = i) and then compare these values. The working parameters (nm, ) were
determined when the map was at its highest efficiency for the corresponding working point. By two-degree-offreedom control, the interaction between electric motor efficiency and hydraulic pump efficiency could beconsidered properly. Compared to one-degree-of-freedom control, the proposed control method is controlled by
two control parameters (nm and ) so it is able to achieve high efficiency for a given working point. Moreover,the flexibility of nm and  could help the system operate within its permitted working range (motor speed nm andtorque Tm, hydraulic pump speed np and pressure p) and thus avoid overloading in the servo motor andhydraulic pump.

EXPERIMENTAL VALIDATION
Experiment for establishing overall efficiency maps
To validate the simulation results, the experiment is conducted in EHDS included servo motor and hydraulicaxial piston pump as shown in Figure 5. The working points for experiment of servo system and hydraulicpiston pump shown in Figure 6 and 7, respectively. The working points are measured at 09 displacement ratio = 0.2, 0.3, 0.4, 0.5 0.6, 0.7, 0.8, 0.9, 1.0.

Motor torque Tm (Nm)Pressure differential p (MPa)Figure 5. Schematic of hydraulic circuit for measuring the efficiency of system
1. Servo Amplifier; 2.AC Servomotor; 3. Rotation Speed Sensor; 4. Torque sensor; 5. Variable Displacement Pump;
6. Pressure Gauge; 7. Pressure Sensor; 8. Flow meter; 9. Relief Valve; 10. Oil Tank.

Motor speed nm (rpm)Figure 6. Scatter chart of experiment of servo
motor

Pump speed np= nm (rpm)Figure 7. Scatter chart of experiment of hydraulicpiston pump

From measured data, experimental coefficients in theoretical formulas (1) and (2) are calculated by the leastsquares method. Then experimental efficiency maps of servo motor and hydraulic pump are created fromformulas with experimental coefficients to build 09 overall efficiency maps of EHDS at 09 displacement ratio .
To validate the advantage of proposed control method, these overall efficiency maps are used to determine
overall efficiency of two-degree-of-freedom control (VS-VP) and one-degree-of-freedom control (VS-FP, FSVP). The efficiency comparison of these system is shown in Figure 8.
In Figure 8, VS-VP unit achieved better overall efficiency than others at almost all given working points
because of the flexibility of both nm and . The advantage of the VS-VP unit can be seen clearly at a low flowrate range (Q = 210 L/min). Considering the work point of Q = 2 L/min; p = 16 MPa, the efficiency of the3

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VS- VP unit is 28% and 10% higher than that of the VS-FP unit and FS-VP unit, respectively. These results in
Figure 8 matches with the simulation results as shown in authors’ previous research.

Improving the transient efficiency of EHDS using efficiency maps

Overall efficiency tOverall efficiency t

The results shown in Figure 8 demonstrated the advantage of proposed control strategy in separated working
points of EHDS. For applying in actual EHDS, the controller should improve the overall transient efficiency ofthe system continuously. By using overall efficiency maps, the controller could select the high-efficient working
point at any required flow rate Q and pressure differential p. The Figure 9 shows the flow rate command withstep signal and Figure 10 shows overall efficiency comparison of 3 types of EHDS controls (VS-VP, VS-FP,
FS-VP). The results shown that the efficiency of VS-VP system is always higher than VS-FP and FS-VP system.
This means that the EHDS with two-degree-of-freedom control could achieve higher efficiency than onedegree-of-freedom control.

Flow rate Q (L/min)
Overall efficiency tOverall efficiency tFlow rate Q (L/min)

Flow rate Q (L/min)

Flow rate Q (L/min)
Flow rate Q (L/min)Figure 8. Overall efficiency t comparison at pressure differential p = 16, 10, 6.3, and 4 MPa

Time
Figure 9. Step flow rate command Q (L/min) at constant pressure p = 4 MPa

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Overall efficiency t

The 11th JFPS International Symposium on Fluid Power 2020

Time
Figure 10. Overall efficiency t comparison at output different pressure p = 4 MPa

CONCLUSION
This study conducted experiment to validate the simulation results in previous research of authors. The
experimental shown that two-degree-of-freedom control could achieve higher efficiency than one-degree-offreedom control not only at separated working points but also entire operating range with step flow ratecommand. In next stage, this control will be set up and validated in actual machines.

REFERENCES
1. WILLKOMM, J., et al., Quadratic Programming to Optimize Energy Efficiency of Speed- and Displacement-Variable
Pumps, Proceedings of the 8th FPNI Ph.D Symposium on Fluid Power, Lappeenranta, Finland, 2014, V001T05A002,
pp.1-10.
2. PHAN, H.T., SATO, Y., Improving the Overall Efficiency of an Electro-Hydraulic Drive System by using Efficiency
Maps, JFPS International Journal of Fluid Power System, 2021, Vol.14, Issue 1, pp.10-18.
3. THE JAPAN HYDRAULICS & PNEUMATICS SOCIETY, Oil Hydraulics and Pneumatics Handbook, Ohmsha, Ltd.,
1989, pp.205-206. (in Japanese)
4. ITAGAKI, S., PHAN, H.T., SATO, Y., Study on High Efficiency Power Transmission in Servo Motor-Driven Hydraulic
System, The Proceedings on Spring Conference of Japan Fluid Power System Society, 2019, pp.45-47. (in Japanese)5

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Genaral Session | Pneumatics

Medical and Welfare[GS3-01] Development of Silicone Outer Shell Type Pneumatic Soft Actuator〇Yuma Nakanishi1, Yasuhiro Hayakawa1, Keisuke Kida1, Hiroaki Ichii1 （1.National Institute ofTechnology (Kosen) , Nara College.）[GS3-02] Development of Pneumatically Driven Verification System for Ophthalmic
Needling Operation
Feng Tao1, 〇Maina Sogabe2, Taro Ito3, Tetsuro Miyazaki2, Toshihiro Kawase4,1, Takahiro Kanno5,
Yoshikazu Nakajima1, Kenji Kawashima2 （1.Tokyo Medical and Dental University, 2.The
University of Tokyo, 3.Tottori University, 4.Tokyo Institute of Technology, 5.RIVERFIELD Inc）

[GS3-03] Development of a Whole Body Training Device by Multi-directional Force
Input Using Pneumatic Artificial Muscles
Soichiro Ito1, 〇Tetsuro Miyazaki2, Junya Aizawa3, Toshihiro Kawase1,4, Maina Sogabe2, Takahiro
Kanno5, Yoshikazu Nakajima1, Kenji Kawashima2 （1.Tokyo Medical and Dental University, 2.The
University of Tokyo, 3.Juntendo University, 4.Tokyo Institute of Technology, 5.RIVERFIELD Inc.）[GS3-04] Development of Robotic Forceps Driven by Soft Actuator with Built-In
Displacement Sensor〇Osamu Azami1, Takahiro Kanno1, Toshihiro Kawase2,3, Maina Sogabe4, Tetsuro Miyazaki4, Kenji
Kawashima4 （1.Riverfield Inc., 2.Tokyo Medical and Dental University, 3.Tokyo Institute ofTechnology, 4.Tokyo University）©The Japan Fluid Power System SocietyGS3-01

The 11th JFPS International Symposium on Fluid Power 2020

Development of Silicone Outer Shell Type Pneumatic Soft Actuator
Yasuhiro HAYAKAWA*, Keisuke KIDA*, Yuma NAKANISHI* and Hiroaki ICHII**Department of Control Engineering
National Institute of Technology, Nara College
22 Yata-cho, Yamato-Koriyama, Nara, 639-1080 Japan(E-mail: hayakawa@ctrl.nara-k.ac.jp)
Abstract. The pneumatic soft actuators have both passive flexibility of the casing and active flexibility throughpressurizing and depressurizing. These features are expected to be human-friendly actuators because they enable theapproach of using an original soft system by controlling the internal air pressure. The current pneumatic soft actuator
has two technical points of interest; control method of driving direction and operating pressure level, requiring an
actuator to match these points. In this study, a pneumatic soft actuator with a silicone rubber case, SCSRA (SpongeCore Soft Rubber Actuator) was proposed and developed to respond to the requirement. Thought some experimental
results, it is cleared that SCSRA controlled the driving system by realizing a large stroke in the low-pressure regionof 30kPa or less as well as peeling and adhering the silicone film.
Keywords: High Performance Insole, Pneumatic, Care Prevention, Walking, Soft Rubber Actuator

INTRODUCTION
There are many types of actuators used in today's industrial, medical, and welfare fields, depending on theiroperating principles. Even for a simple piston-cylinder system, there are thermal, electric, and fluid types, andthey are used in different ways depending on the application and operating object. Among them, pneumatic soft
actuators are suitable for applications involving contact with flexible objects such as fruits and human bodies.
These actuators, driven by air and mainly using rubber as casing, are referred to as "human-friendly actuators”.
This is because pneumatic soft actuators have "two flexibility features[1]" that set them apart from otheractuators. First, the flexibility of the case and the compressibility of the air allow the actuator to exhibit passive
flexibility as an elastic body even when it is not pressurized. Secondly, the flexibility can be actively controlled
by changing the shape and stiffness of the actuator by pressurizing inside of it. This feature makes it possible torealize a unique process in which an original soft system can be strengthened and utilized by control.
Furthermore, the system does not require high current or high heat for operation, and even if the case isdamaged, the operating energy is harmless air, resulting in minimal impact on the surrounding environment or
the object being operated. Because of these features, systems using pneumatic soft actuators have beendeveloped and realized in various fields beyond the existing applications in recent years.
There are two major technical points of interest in the pneumatic soft actuators that have been developed to date.
The first is the pressure level to be used. For example, the PBA[2] (Pneumatic Balloon Actuator) performing abending motion requires an internal pressure of 70kPa for a 7mm-motion stroke, and the twist actuator[3]performing a twist motion, requires an internal pressure of 178kPa for a 70°motion. In the case of pneumatic
soft actuators, which are often used in contact with flexible objects, there is a close relationship between theinternal pressure and the safety of the operating energy, so from the viewpoint of the impact on the operating
object and the surrounding environment, driving at even lower pressure is required. The second is the structure
to control the driving direction. In order to generate deformation in a specific direction, the pneumatic rubber
artificial muscle[4] uses a sleeve woven with fibers, and the FMA[5] (Flexible Micro Actuator), constrains thedeformation direction by installing multiple air chambers. If the drive direction can be controlled with a simpler
structure, it will be possible to reduce the cost and space of the actuator.
This study focuses on these improvements and aims to develop a pneumatic soft actuator, SCSRA (Sponge Core
Soft Rubber Actuator), that satisfies the following three requirements.
(1) Low-pressure operation
A large internal pressure to motion stroke ratio can be achieved at lower pressure levels(small less than30kPa) than conventional pneumatic soft actuators.
(2) Simple control method for driving direction
By keeping the pressure level low, the deformation direction can be constrained at low cost and with asimple structure.
(3) Both passive and active functions
It is equipped with both pressurized and non-pressurized functions by taking advantage of the "two-1

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flexibility" characteristic of pneumatic soft actuators.
In this paper, the structure and fabrication method of the SCSRA were described first, and the functions of the
SCSRA to satisfy three requirements mentioned above are explained. Next, the characterization experiments
show that the specified functions could be achieved under both pressurized and unpressurized conditions, andthe parameters affecting the characteristics are clarified.

SCSRA(SPONGE CORE SOFT RUBBER ACTUATOR)

Structure and Features
The structure of SCSRA is shown in Figure1. SCSRA is a pneumatic soft actuator consisting of a core foam
sponge surrounded by a case made of silicone rubber and a polyurethane tube for inflow and outflow of air. The
combination of three flexible materials and air pressure is suitable for applications involving contact withflexible objects, such as fruits and human bodies.
The difference from conventional pneumatic soft actuators is that the pressurized area is not just a cavity, but a
gap between the continuous foam sponges. The presence of the sponge enables the actuator to maintain a certain
level of rigidity even when the inside of the actuator is open to the atmosphere, and it also has the advantage thatthe same internal pressure can be obtained at a lower flow rate than when the inside of the case is simply
pressurized.
Since the case is made of two-component RTV rubber, which is liquid before curing, any actuator shape andpressure area can be achieved by changing the shape of the sponge and mold. In addition, the sponge andsilicone rubber may be adhered or separated during fabrication, so the functions described in the next section
can be used.

Figure 1. Structure of SCSRA

Fabrication procedure
Figure 2 shows the fabrication procedure of the SCSRA. The procedure is roughly divided into (1)fabrication of
the bottom surface, (2)processing and placement of the sponge, (3)fabrication of the side and top surfaces, and(4)installation of the tube. The bonding conditions between the silicone rubber and sponge may vary dependingon the timing of the sponge placement. Specifically, in the case of adhesion, the sponge is placed before thesilicone liquid on the bottom becoming elastic, and in the case of separation, the sponge is placed after the liquid
becoming elastic.
For convenience, SCSRAs in which silicone rubber and sponge are adhered are referred to as "adhesive
elements," and SCSRAs in which silicone rubber and sponge are separated are referred to as "separated
elements”.

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Figure 2. Manufacturing Procedures of SCSRAFunctions
The SCSRA has the following four functions, depending on whether the inside of the actuator is pressurized or
unpressurized, and whether the sponge and silicone rubber are adhered or separated.
(1)Force-distribution function (Figure 3(a))
The passive function of the SCSRA is demonstrated by sealing the inside of the actuator with atmosphericpressure. The flexibility of the constituent materials themselves and the compressibility of the air disperse theforce applied from the outside in multiple directions.
(2)Force-estimation function (Figure 3(b))
The passive function of SCSRA is demonstrated by sealing atmospheric pressure inside the actuator andconnecting a pneumatic sensor to the tube. When an external force is applied to the SCSRA top surface, thepressure-receiving area and the indicated value of the air pressure sensor are substituted into Equation (1) toestimate the applied external force.
P=F/A

(1)

P : Pressure[Pa]F : Vertical force acting on a surface[N]A : Force-activated area[m²](3)Rigidity-changing function (Figure 3(c))
The active function of SCSRA is demonstrated by fully adhering the sponge and silicone rubber, andpressurizing the inside of the actuator. Deformation of the silicone membrane under pressure is constrained by
the sponge, which rapidly increases its stiffness in response to slight changes in shape.
(4)Shape-changing function (Figure 3(d))
The active function of SCSRA is demonstrated by separating the sponge and silicone rubber only on a specificsurface and pressurizing the actuator. Since the silicone rubber other than the separating surface is constrainedby the sponge, only the separating surface is deformed significantly under pressure. The driving direction can becontrolled by changing the separation surface.

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The 11th JFPS International Symposium on Fluid Power 2020(a) Force-distribution Function(b) Force-estimation Function(c) Rigidity-changing Function(d) Shape-changing Function
Figure 3. Functions of SCSRACHARACTERIZATION EXPERIMENTS
Objective
In order to show that SCSRA satisfies the three requirements described in Chapter 1, characterization
experiments were conducted. As explained in Section 2.2, there is a close relationship between the threerequirements and the functions of SCSRA. Therefore, by showing that SCSRA can realize the four functions,the purpose of this study is considered to have been achieved.
The function (2), the external force estimation function, has been realized in the previous research in thislaboratory, the high-performance shoes system for gait training[6] (Figure 4). The SCSRA is used as an insole
element and realizes the function of understanding the subject's gait state through the foot pressure distribution.
In this study, the following functions are demonstrated through characteristic experiments: (1) external force
dispersion function, (3) stiffness change function, and (4) shape change function.

Figure 4. Insole Elements of High-performance Shoes

Use of Devices
Table 1 shows the devices are used for the characteristic experiments.

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Table 1 List of Equipment Used in the Experiments
Device Name
DC Stabilized Power Supply
Control Circuit
AD Converter
Laser Displacement Meter
Micro Pump
Pressure Sensor
3-Port Spool Valve

Manufacturer
Gwinstek
Gwinstek
METRONIX
ArduinoELMOSFASTUS
ParkerFujikura
SMC

Model Number
GPS-3030D
GPS-3030D
512C
UNO R3
RAI-16
CD22-35V
CTS
AG213-200KG
S070C-RDC-32

Experimental Devices
Figure 5 shows the experimental apparatus used in the characterization experiments. The system is divided into
two parts: one is to control the internal pressure of the SCSRA, and the other is to measure the displacement
caused by pressurization or load application using a laser displacement meter. The final output value is obtainedby converting the analog output of the laser displacement meter to a digital value using an A/D converter. Theinternal pressure of the actuator is controlled by the electro-pneumatic regulator, and the output of the airpressure sensor is obtained as a voltage value. The breakdown of V1~V7 shown in Figure 5 is as follows.
V1~V3：Supplied by a DC stabilized power supply.
V4~V6：Supplied from each pin of the Arduino, the control circuit.
V7：Supplied from the USB port of the PC.

Figure 5. Experiment Equipment for Experiments

SCSRA for the Experiments
The parameters affecting the properties of SCSRA are shown in Figure 6, and the physical properties of the
sponge and silicone rubber used in SCSRA are shown in Table 2 and Table 3. In this study, the SCSRA with theparameters shown in Table 4 is used as the reference SCSRA, and we examined how the overall characteristics
changed when each parameter is changed.

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Figure 6. Main Parameters of SCSRA

SpongeDensity [kg/m3]
Hardness [N]Table 2 Main Property of Sponge
Stiff Urethane Sponge EMO
50 5
12.75 [kPa]

Silicone

Table 3 Main Property of Silicone
KE-1316

Hardness [Durometer A]

23

Viscosity [Pa s]

35

Tensile Strength [MPa]6.5

Elongation at Break [%]700

Tear Strength [kN/m]

33

Line Shrinkage Ratio [%]0.1

Density [g/ cm3]

1.13
Table 4 Parameters of Reference Actuator
50×50×9

External Shape [mm]Wall Film Thickness [mm]Upper and Lower Film
Thickness[mm]Internal Pressure P [kPa]
Adhesive conditions on the Top
Surface420Separated

EXPANSION CHARACTERISTICS EXPERIMENT
Experimental method
The following procedures are used for the experiment.
(1) Using the experimental setup shown in Figure 5, the internal pressure of the reference SCSRA is set to 0, 5,
and 10 kPa.
(2) The relationship between the internal pressure and the amount of expansion is obtained by measuring the
top of the expansion shape (the point at which the amount of expansion is maximum) using a laser
displacement meter.
From the standard SCSRA, only the bonding condition of the top surface is changed, and the same procedure is
performed for the SCSRA in which the silicone film and sponge has completely adhered.

Results and Discussion
The relationship between the actuator internal pressure p and the amount of expansion for the two types of
SCSRA is shown in Figure 7. The results can be summarized as follows.
(1) The expansion characteristics of SCSRA followed a quadratic curve:

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The correlation coefficients for the quadratic curve approximations were above 0.9 for both the separated
and adhered elements, indicating a very strong positive correlation. This is thought to be largely due torubber elasticity, which is a key property of silicone rubber. From the characteristic curve, it can be seenthat the increasing rate of the expansion decreased with raising internal pressure, especially for thereference SCSRA.
(2) The characteristics of separated and adhered elements could be clearly distinguished:
The property curves show that there is a clear difference between the properties of the separated andadhered elements. The first-order coefficient, which is the most important term in the approximation
equation, shows a difference of about 24.9 times, and the expansion amount at an internal pressure of 15
kPa shows a difference of about 38.2 times, indicating that the rigidly- and shape-changing functions of
SCSRA could be switched depending on the adhesion conditions of the silicone film.
(3) The separate elements had a high internal pressure-to-operating stroke ratio:
Focusing on the reference SCSRA in the expansion characteristic curve, an expansion of approximately
26.0 mm was achieved at an internal pressure of 15 kPa. This indicates that the SCSRA had a shapechanging function. Although it is difficult to compare simply with the bending and twisting motions in theprevious studies described in Chapter 1, the internal pressures used were 4.6 and 11.9 times those of the
SCSRA, respectively, and the motion stroke was equal to or less than that of the SCSRA. This indicatesthat the SCSRA achieved higher internal pressure and motion stroke than conventional pneumatic soft
actuators.

Figure 7. Expansion Amount Characteristic of SCSRASPRING STIFFNESS CHARACTERISTICS EXPERIMENT
Experimental method
The following procedures are used for the experiment.
(1) Using the experimental setup shown in Figure 5, the internal pressure p of the reference SCSRA is set to10, 20, and 30 kPa.
(2) A load of 100, 200, 300, 400, 500, 1000, and 2000 g is applied to the SCSRA top surface, and the amountof sinking is measured by a laser displacement meter to obtain the spring stiffness characteristics.
(3) The internal pressure of the reference SCSRA is set to 0 kPa, and the wall thickness a (Figure 6) is changed
to 2, 4, and 6 mm, respectively, and the same measurements are made using the procedure (1)~(2).
(4) The internal pressure of the reference SCSRA is set to 0 kPa, and the upper and lower film thicknesses, b,
are changed to 2 and 4 mm, respectively, and the same measurements are made using the procedure (1)~(2).

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Results and Discussion
Figure 8 shows the stress-strain curves under different internal pressure “p” from the reference SCSRA. Notethat only 100, 200, 300, 400, and 500 g loads are applied to the data because excessive stress under pressure
may cause the actuator to break. The results are as follows.
(1) The spring stiffness characteristics of SCSRA followed an exponential function:
In common with each “p”, the correlation coefficient for the exponential function approximation exceeded
0.9, indicating a very strong positive correlation. This characteristic, in which the required stress increased
with raising displacement, was extremely similar to the characteristics of air and rubber materials during
compression, suggesting that silicone rubber and compressed air formed the spring stiffness characteristics.
The effect of the sponge inside wasn't investigated in this experiment, but had been detailed in a previous
study conducted by our laboratory[7]. In the experiment using two types of silicone rubber and five typesof sponge, it was pointed out that the spring stiffness characteristics changed depending on the
combination of silicone rubber and sponge.
(2) The spring stiffness increased with raising internal pressure:As the “p” increases, the stress required to generate a certain amount of displacement also increases,
indicating an increase in spring stiffness. It can be said that the SCSRA realized the rigidly-changing
function with the results of extended characterization experiments.

When the initial internal pressure was applied, small deformation was observed as shown in Table 5. Since thedeformation was less than 3% of the total SCSRA thickness of 9 mm, it was not considered to be a problem inthe experiment.

Figure 8. Stiffness Characteristic of SCSRA by P(10.20.30kPa)
Table 5 Initial expansion by pressurization
Initial Internal Preassure “p” [kPa]
Displacement [mm]
10
0.02
20
0.08
30
0.21

Spring stiffness characteristics by wall film thickness A
The stress-strain curves for wall film thickness “a”(Figure 6) different from the reference SCSRA are shown in
Figure 9. The results show that as “a” increases, the stiffness of the entire SCSRA decreases. This may be due tothe fact that the stiffness of the silicone rubber used in the experiment is less than that of the sponge. In addition,even if the interior of the SCSRA is measured at atmospheric pressure, the characteristics follow an exponentialfunction as in the case of pressurization. This indicates that the SCSRA has the capability to distribute external
force due to the passive flexibility of silicone rubber, sponge, and air under non-pressurized conditions.

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Figure 9. Stiffness Characteristic of SCSRA by A(2,4,6mm)

Spring stiffness characteristics by upper and lower film thickness B
Figure 10 shows that the stress-strain curves for the reference SCSRA and upper and lower film thicknesses“b”(Figure 6). The results show that as “b” increases, the stiffness of the entire SCSRA decreases. This can be
attributed to the decrease in the spring stiffness according to the series connection equation (Figure 11) because
the increase of the top and bottom film thicknesses corresponds to the addition of a silicone film in series withthe reference SCSRA.

Figure 10. Stiffness Characteristic of SCSRA by B(2,4mm)

Synthetic spring constant k

Figure 11. Rigidity Spring Constant in Series Connection

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CONCLUSIONSIn this study, a pneumatic soft actuator, SCSRA, with a sponge core and a silicone rubber case was developed.
First, three requirements for SCSRA were defined based on advantages and problems of the previous studies,and then the requirements were specified by passive and active functions. Then, we showed that the SCSRA canrealize these functions through characterization experiments.

ACKNOWLEDGMENTS
I would like to express my sincere gratitude to Associate Professor Pak Keunyoung for her english guidance
during her busy schedule in conducting this research. I would also like to thank SMC Corporation for providing
some of the experimental equipment. Finally, I would like to express my utmost gratitude to all the students who
supported my research, both inside and outside the laboratory.

REFERENCES
1. T. Noritugu., Pneumatic Soft Actuator and Human Friendly Mechanism, Journal of Robotics and Mechatronics, Vol.21,
No.7, pp. 722-726, 2003
2. Benjamin Gorissen., Michaël De Volder., et al., Theoretical and experimental analysis of pneumatic balloon
microactuators, Sensors and Actuators A, Vol.168, pp. 58-65, 2011
3. B.Gorissen., T.Chiskiro., et al., Flexible pneumatic twisting actuators and their application to tilting micromirror，Sensors and Actuators A, Vol. 216, pp. 426-431, 2014
4. S. Kawamura., M. Nokata., et al., Control actuator basics, Corona publishing co., ltd., 2014, pp 49-50
5. K. Suzumori., Flexible Microactuator, Transactions of the Japan society of mechanical engineers. C, Vol. 55, No. 518,
pp.2574-2552, 1989
6. Y. Hayakawa., Y. Kimata., et al., Study on Human Behavior Classification by using High-Performance Shoes Equippedwith Pneumatic Actuators, Journal of Robotics and Mechatronics. Vol.32, No.5, 2020 pp. 1-9
7. T. Minamoto., Y. Hayakawa., Development of a New Rubber Element with Human Compatibility, Graduation Researchof Department of Control Engineering, NIT, Nara College. 2020

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Development of Pneumatically Driven Verification System
for Ophthalmic Needling Operation
Feng Tao*, Maina Sogabe**, Norihiko Ito***, Tetsuro Miyazaki**, Toshihiro Kawase*, ****,
Takahiro Kanno***** and Yoshikazu NAKAJIMA*, Kenji Kawashima ***Institute of Biomaterials and Bioengineering, Tokyo Medical and Dental University
2-3-10 Kanda Surugadai, Chiyoda-ku, Tokyo, 101-0062 Japan**
Department of Information Physics and Computing, The University of Tokyo
7-3-1 Hongo, Bunkyo-ku, Tokyo, 113-8656 Japan(E-mail: kenji_kawashima@ipc.i.u-tokyo.ac.jp)**
Department of Agriculture, Tottori University
4-101 Koyama-cho Minami, Tottori, 680-8550 Japan****
Institute of Innovative Research, Tokyo Institute of Technology
4259 Nagatsuta-cho, Midori-ku, Yokohama, Kanagawa, 226-8503 Japan*****
RIVERFIELD Inc., Yotsuya Medical Bldg. 5th floor, 20 Samon-cho, Shinjuku-ku, Tokyo, 160-0017 Japan

Abstract. The demand for surgical assist robots is increasing in the field of ophthalmic surgery. However, the patient's
movements that occur during surgery increase the difficulty of the operation, and the need for a system that can
reproduce and verify the movements has arisen. The proposed system is mainly composed with two parts, needlingmechanism and head rotation mechanism. The needling part consist of a remote center of motion mechanism for tilt
and roll movements and insertion. The head rotation part is aimed to replicate the posture of the patient eyeball andhead. In order to reproduce the smooth motion like a living body, we introduced a pneumatic cylinder with low friction
and back drivability. In this study, we used the proposed system to determine the needling point and verified its
accuracy. In the future, we plan to use this device to further evaluate the needling task for ophthalmic surgery.
Keywords: Ophthalmic surgery, Surgical assist robot.

INTRODUCTION
The demand for eye surgery increases year by year, and cataract surgery alone exceeds 1 million cases annually[1, 2]. In order to improve the accuracy and prognosis of these operations, research on surgery assisted robotic
systems is actively conducted. Surgery assisted robotic systems have many advantages, including enhanced
accuracy dexterity, elimination of tremors, capacity for telesurgery, task automation, a rapid learning curve, andalleviating the surgeon’s burden [3-6]. Among them, the issue is the development of surgical techniques thatmatch the body movements of patients during surgery. At the time of surgery, the patient can move the eyeball tosome extent with partial anesthesia; it is necessary to have the patient face the front.
However, some patients move their eyes and head due to repellent reactions (Fig. 1). In ophthalmic surgery, suchas retinal vein occlusion and cataract, a needle may be inserted into the eye to recanalize the thromboembolic siteor to inject drugs [7]. People in the medical field use robust fixation and sedatives to prevent unpredictable
movements as much as possible. However, some patients are afraid of being fixed or are resistant to sedatives andmay not be able to fix them. Currently, mechanical assist tools in ophthalmic surgery assume that the patient islooking straight ahead, so they do not assume that the head is rotated. Therefore, our goal was to perform theneedling operation on the sclera at the corneal edge, which is often performed in ophthalmic surgery, under theeye movement when the head is rotated.
In this case, it is important to have smooth movements that are typical of living bodies. To reproduce this, wechose a pneumatic cylinder, which has less friction and back-drivability than conventional robots using electricmotors. In our previous research, we have developed a device that automates the needling operation [8, 9], and as
the next step, we aimed at the insertion point of the needle into the eyeball in accordance with the movement.
Therefore, we developed a pneumatically driven device that allows us to check the needling operation under thehead movement.

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Figure 1. Background about needling for ophthalmic surgery.

CONTROL SYSTEM OF THE PROPOSED SYSTEM
The requirements for this system are that there should be low friction during needling operation, which wouldinterfere with precise movement, and that safety should be ensured when the system is used on a human body inthe future. For this reason, we chose a pneumatic drive system that has low friction during operation and also hasback drivability. Fig. 2 shows the overview of the proposed system which is mainly composed with two parts,
needling mechanism and head rotation mechanism. The needling system consists of two pneumatic cylinders forvertical and horizontal coordinate movement, two pneumatic rotary actuators for pitch and roll rotation, and one
soft pneumatic actuator for the insertion of the needle. The head rotation system consists of two pneumatic
cylinders for vertical and horizontal coordinate movement and one pneumatic rotary actuator for head tilt.
As shown in the Fig. 3a, the proposed system is actuated by the air pressure. Compressed air supplied fromcompressor (supply pressure: 0.4 MPa) passes through the regulator and is supplied to the 5-port servo valve(FESTO MPYE- 5-M5-010-B). Flow rate of servo valve depending on its input voltage, which can be controlled
by the control PC. The air supplied from the servo valve will drive the actuators used in needling mechanism andhead rotation mechanism. Through AD converter, data of position sensors and pressure sensors can be read andcontrol signal will be transferred to the servo valves through DA converter.
Fig. 3 b shows the diagram of real time position control system. In this system, a cascade control system composedwith a main loop (position control) and minor loop (pressure control) was adopted. Cascade control has strength
in the suppressing of the influence from disturbance and can achieve a high-level of position control. Parametersof cascade control are shown in the Table 1 and 2, where Kpp is the proportional gain of position control, Kpi is theintegrated gain of position control, Kap is the proportional gain of pressure control, Kai is the integrated gain ofpressure control, and u is the input voltage of servo valve.

Figure 2. Overview of the proposed system.

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(a)

(b)

Figure 3. Diagram (a) and block diagram (b) of experimental setting.

Actuator #
1
2
3
4
5

Actuator #67
8

Table 1. Parameters of PID gains in needling system
Kpp
Kpi
Kap
2.5 N/mm
45.0 kPa/mm
37.5 kPa/mm
35.0 kPa/mm
35.0 kPa/mm

1.5 N/(mm・s)
25.0 kPa/(mm・s)
35.0 kPa/(mm・s)
15.0 kPa/(mm・s)
20.0 kPa/(mm・s)

1.5 V/N
0.03 V/kPa
0.03 V/kPa
0.03 V/kPa
0.03 V/kPa

Table 2. Parameters of PID gains in head rotation system
Kpp
Kpi
Kap]
28. 0N/mm
35.0 kPa/mm
25.0 kPa/mm

15.0N/(mm・s)
20.0 kPa/(mm・s)
12.5 kPa/(mm・s)

1.5 V/N
0.03 V/kPa
0.03 V/kPa

Kai
2.5 V/N・s
0.05 V/(KPa・s)
0.05 V/(KPa・s)
0.05 V/(KPa・s)
0.05 V/(KPa・s)

Kai
0.25 V/(N・s)0.05 V/(KPa・s)
0.05 V/(KPa・s)

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EXPERIMENTAL PROCEDURE AND RESULT
Needling System

Figure 4. Overview of the needling system.

In the manufacturing of needling mechanism, w two rotary actuators (KURODA PRNA3S-180-90) and two linear
actuators (SMC CY3R10-60) are employed for pitch and roll degrees of freedom, and one linear actuator (SMCCJ2XB10- 45Z) is employed for Z-direction (depth) movement. 45Z) for Z-direction (depth) movement. Under acondition of no risk of vision interference and collision with optical device, the needling mechanism is designed
and manufactured to realize a 20 degree of workspace in the freedom of pitch and 35 degree of workspace in thefreedom of roll. The prototype of needling mechanism is shown as Fig. 4.
According to the data in Fig. 5, we can find that the tip of end effector can move around remote center of motion(RCM) with a range of accuracy error. To the detail, in the freedom of pitch with a motion range of 20 degree, theerror in x direction is from -0.25mm to 1mm, the error in z direction is from -2mm to 1.5mm. In the freedom ofroll with a motion range of 32 degree, the error in y direction is from -1.75mm to 1.75 mm, the error in z direction
is from -1.5mm to 0.25mm. The accuracy error of manipulation mainly comes from the defects in the mechanicaldesign and control system. As for the mechanical design, manufacturing error of the prototype, friction of actuator(especially rod-less linear actuator need to overcome a large friction during manipulation), and the unbalanced
gravity effect of needle holder to the actuator 4 and actuator 5 will also influence the accuracy of manipulation.
As for the control system, firstly, the accuracy of RCM mechanism adopted (instantaneous type) heavily relies on
the control performance of each actuator used in the manipulation (compared with the Parallelogram-based type)
and the accumulation of error from each single actuator reduce the accuracy error of the needling mechanism.
Secondly, there is an improvement in the parameter tuning of PID control, which will lower the delay in the
maximum/minimum value of motion range.

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Figure 5. Evaluation experiment of needling mechanism: the error of RCM in the freedom the pitching (a,b,c) and the roll(d,e,f) motion experiments.

Head Rotation System

Figure 6. Overview of the head rotation system.

To design the proposed head rotation mechanism, we need to discuss the human head motion and eyeball motion
during the operation in the following sections. The motion of human head is based on the motion of cervical spine,which can be divided into three types: Flexion/Extension, Rotation and Lateral Rotation.
Fig. 6 shows the overview of the head rotation system part. Head rotation mechanism is designed to realize a 20degree head rotation in with Y axis and compensate the associated motion of eyeball center during rotation. In the
manufacturing of head rotation mechanism, we adopted one rotary actuator (KURODA PRNA3S-270-45) andtwo linear actuators (SMC CY3R10-60) to realize the workspace mentioned above. Furthermore, to confirm thefeasibility of the proposed robotic system, we developed a human eyeball phantom used for the evaluation
experiment. The size of this phantom is mainly based on a commercial skull model.

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Figure 7. Evaluation experiment; Error of compensation during head rotation.

The control accuracy of head compensation in the process of head rotation, whose performance will decide that if
the patient’s eyeball keep in the detectable range of other imaging device. The data of this experiment is to measure
the error between reference value and measured value of compensation in x and z direction.
Based on the data shown in Fig.7, it is confirmed that the proposed system will have a head compensation ability
in x and z direction with a range of accuracy error during a rotation range of 20 degree. To the detail, the error inx direction is from -2.3 to 2mm, the error in z direction is from -1.5mm to 2.0mm. Under this range of error,
manipulator can keep the human eyeball in the detective range of observation. However, on the other hand, theerror of head compensation is related to the accuracy of needling operation, this is because, kinematics of needlingmechanism is based on a reference value of compensation. Therefore, this kind of error should be considered in
the improvement of kinematics in the next step.

CONCLUSION
We developed an eye-phantom device that can reproduce the time of head rotation considering the human skeleton
and verified the accuracy of the actual needling operation using the proposed system. As a result, it was foundthat the head rotation could be reproduced using the pneumatic driven proposed system, and the needling operation
accuracy could be evaluated. The error of the insertion point setting for needling operation was 2 mm. The errorof the head rotation mechanism was up to 2.3mm. Regarding the error, competition for reference bulbs on theneedle system side is considered, and it is suggested that it can be solved by improving them.
In future, we plan to improve needing accuracy in real-time using the proposed device and to examine theadaptability of manipulation using an enucleated eyeball instead of the phantom.

ACKNOWLEDGMENTS
This work was supported in part by Riverfield Inc. .

REFERENCES
1. SHARMA, A., D’AMICO, D. J., Medical and surgical management of central retinal vein occlusion. International
Ophthalmology Clinics, 2004, 44(1): 1-16.
2. ROGERS, S. et al., The prevalence of retinal vein occlusion: pooled data from population studies from the United States,
Europe, Asia, and Australia. Ophthalmology, 2010, 117(2): 313-319.
3. JEGANATHAN, V., SWETHA, E. and SHAH, S., Robotic technology in ophthalmic surgery. Current opinion in
ophthalmology, 2009, 21. 75-80.
4. Computer Motion., Available at http://www.computermotion.com. Accessed 2003.
5. DAKIN G. F., GAGNER M., Comparison of laparoscopic skills performance be- tween standard instruments and two
surgical robotic systems [abstract]. Accepted for the 8th World Congress of SAGES, New York, NY 2002.
6. HIMPENS J., LEMAN G., CARDIERE G. B., Telesurgical laparoscopic cholecystectomy. Surg Endosc 2003;17:574–9.
7. ASAMI, T. et al., An Endovascular Cannulation Needle with an Internal Wire for the Fragmentation of Thrombi in Retinal
Vein Occlusion. Transl Vis Sci Technol. 2016;5(5):9.
8. KANNO T. et al., Soft-material Robot for Ophthalmologic Surgery. Sensors and Materials, 2020 32. 1041.
9. TAO F. et al., Development of a robotic platform for OCT guided retinal surgery. The 38th annual conference of the
Robotics Society of Japan (RSJ), 2020.

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Development of a Whole Body Training Device
by Multi-directional Force Input Using Pneumatic Artificial Muscles
Soichiro ITO*, Tetsuro MIYAZAKI**, Junya AIZAWA***, Toshihiro KAWASE*, ****,
Maina SOGABE**, Takahiro KANNO***** , Yoshikazu NAKAJIMA*
and Kenji KAWASHIMA***

Institute of Biomaterials and Bioengineering, Tokyo Medical and Dental University
2-3-10 Kanda Surugadai, Chiyoda-ku, Tokyo, 101-0062 Japan**
Department of Information Physics and Computing, The University of Tokyo
7-3-1 Hongo, Bunkyo-ku, Tokyo, 113-8656 Japan(E-mail: Tetsuro_Miyazaki@ipc.i.u-tokyo.ac.jp)***
Faculty of Health Science, Juntendo University
2-1-1 Hongo, Bunkyo-ku, Tokyo, 113-8421 Japan****
Institute of Innovative Research, Tokyo Institute of Technology
4259 Nagatsuta-cho, Midori-ku, Yokohama, Kanagawa, 226-8503 Japan*****
RIVERFIELD Inc., Yotsuya Medical Bldg. 5th floor, 20 Samon-cho, Shinjuku-ku, Tokyo, 160-0017 Japan
Abstract. Many of conventional training devices target to train local body parts by repeating simple exercises, andvarious kinds of training menu will be required to train the whole body. In this paper, we propose a novel pneumaticwhole body motion training device by multi-directional force input using pneumatic artificial muscles (PAMs) for
total motor system enhancement. PAM is used to actuate the proposed device and gives a trainee flexible and safe
force. By using the proposed device, efficient training of the whole body motor system will be realized. The first
prototype of the proposed device was developed and applied to a balance board training experiment. In theexperiment, we measured floor reaction force and torque in two cases, (i) training with the PAM force, (ii) training
without the PAM force. As a result of the experiment, we confirmed that the proposed training will be effective toimprove the trainee’s core balance performance in a shorter time compared with the conventional training.
Keywords: whole body motion training device, pneumatic artificial muscle, core balance training.

INTRODUCTION
Background
One of major causes that elderly people become patients who need primary nursing care is motor systemdisorders [1][2], and an appropriate exercise habit is important for realizing a healthy and longevity society. For
example, conventional researches have reported that resistance training can improve muscle performance evenin elder people [3]–[6]. Due to the fitness boom in recent years, training demand for a wide range of age groups
is increasing. In training, a muscular system related to force magnitude and a nervous system related to reaction
speed are trained which are main parts of a motor system [7]. For example, as conventional training methods,
there are muscle strength training using weights or air pressure [8] and nervous system training that reacts
quickly to the external force like a jump training program [9].
Some training devices have the driving source and support specific training which will be difficult by humanpower alone. For example, an electric treadmill is one of the typical training devices with the driving source, andwalking or running training at a target speed set by the trainee is possible by motor control. In addition, various
training devices with the driving source have been proposed in conventional researches. Carignan et al.
developed an arm curl machine that adjusts the load during operation using a motor [10]. In this machine, theload can be adjusted according to the change of the trainee’s elbow angle. The DD System ELITE [11] is a footpress training machine using pneumatic artificial muscle (PAM). The trainee kicks a plate actuated by the PAM
while watching the virtual landscape and the data on the display during the exercise. The electric treadmill
SKILLRUN [12] has a function to automatically adjust inclination and speed during running, and it has also a

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function to train kicking force of leg by controlling the rotational resistance of the running belt. The electrictraining machine series [13] controls the load and allows training both when the muscle contracts and stretches.
These training devices can give the desired load of the trainee by driving the actuator instead of the weight block.
However, the conventional training devices have many categories according to training courses, and themuscular system and the nervous system require individual training menu. Moreover, many of the conventional
devices target to train local body parts by repeating simple exercises, and various kinds of training menu will be
required to train the whole body.

Contribution of This Study
Issues of the conventional training devices include (1) individual training of the muscular system and thenervous system are required, (2) trainee's exercise is limited to simple repetitive motion, and (3) various kinds oftraining are required to train the whole body. In order to solve these issues, we propose a concept of a novel
pneumatic whole-body motion training device by multi-directional force input using PAMs for total motor
system enhancement. The PAM is used to actuate the proposed device and gives the trainee flexible and safe
force. By controlling the contraction force of the PAM and applying an unpredictable load to the trainee, themuscular system and the nervous system are comprehensively trained. By using the proposed device, efficienttraining of the whole-body motor system will be achieved.
Contributions of this study are as follows.
1. The concept of the novel whole body motion training device for total motor system enhancement is
proposed.
2. A first prototype of the proposed device is developed.
3. The prototype device is applied to a balance board training, which is an actual training example. The
performance of the prototype device is verified by an experiment of this training.

WHOLE BODY MOTION TRAINING DEVICE
Concept of the Proposed Device
Fig. 1 shows the concept of the pneumatic whole body motion training device proposed in this study. In Fig. 1, a
rigid frame fixed on the environment and the trainee's body are connected by multiple PAMs. By using the
proposed device and applying multi-directional contraction force of the PAMs to the trainee’s body, the trainee
can train the body balance which is important to keep the posture stable against any external force. In addition, it
is possible to give rotational force by using long PAM wrapped around the trainee's body, and the proposed
device can be also applied to whole body multi axis motion training including torsion, for example, throwing
skill practice of judo, etc. Furthermore, by combining the proposed device with VR glasses, it will be possible toconstruct a whole body motion simulator with force sense presentation function in a virtual space. Informationof the trainee during exercise is measured by several sensors, such as electromyogram, pressure sensor, force
sensor, force plate, motion capture system, etc. The measured information is used for the device control and thetrainee's performance evaluation. As an application example of the proposed device, we mainly assume training
for interpersonal competition with collisions such as team sports and tackle sports.
Novel training method using the proposed device will solve the issues of the conventional training devices suchas (1) individual training of the muscular system and the nervous system are required, (2) trainee's exercise islimited to simple repetitive motion, and (3) various kinds of training are required to train the whole body. About
the issue (1), our research group developed a muscular strength training device actuated by the PAM, andcarried out an experiment to give an unpredictable load to the trainee [14]. As a result of the experiment, when
the unpredictable load was applied, the trainee's muscle activity increased by more than 1.2 times comparedwith a case where the predictable load was applied. By applying the unpredictable load, it will be possible totrain the nervous system related to the reaction speed as well as the muscular strength system. About the issues(2) and (3), the proposed device in Fig. 1 is possible to apply load to the whole body via the soft PAMs without
mechanical limiting the trainee's movement, and the trainee can move spatially and arbitrarily. Therefore, we
consider that the proposed device will be able to solve the issues (1) to (3).

Prototype of the Proposed Device
In this paper, an experimental apparatus as shown in Fig. 2 was developed as the first prototype of thepneumatic whole body motion training device. The prototype device shown in Fig. 2 is composed of analuminum cuboid rigid frame, eight long PAMs, and a force plate. The rigid frame has a width of 1.5 m, a depth

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Figure 1. Concept of the pneumatic wholebody motion training device.

Figure 2. Prototype of the proposed device.

of 1.5 m, and a height of 2.1 m. The trainee stands inside of the frame and its waist is connected to the PAM end,
and another end of the PAM is fixed on the frame. The PAM's contraction force in an arbitrary direction istransmitted to the trainee’s body. The positions of the fixing parts that connect the PAMs to the trainee's bodycan be freely changed according to the exercise task to be performed. In the case of balance board training using
the proposed device, the waist part was selected as the fixing position of the PAMs in order to apply an effective
external force to the trunk. The waist (pelvis) is the center of the body, and when an external force is applied tothis part, it is necessary to move the entire body to balance it. The fixing method of the PAMs is as follows; 1)
wrap a soft supporter around the trainee's waist, 2) attach a nylon belt with fixing parts without slack, 3) connect
the fixing parts and PAMs. This fixing method can firmly apply the PAM contraction force to the trainee's waist
without any pain. The trainee’s ground reaction force during the experiment is measured by the force plate.
When measuring myoelectric potential, the trainee wears electromyogram electrodes. The developed device isthe first prototype, and eight PAMs are used to exert the load on the trainee’s waist. This physical connection
also works as a lifeline and decreases serious accident risks like falling while training. In order to confirm thatthe prototype device can be applied to actual training, a balance board training was selected as an application
example, and we carried out a verification experiment.
The PAM has several unique characteristics. For example, it has a high power-to-weight ratio despite its simple
and lightweight structure consisting only of a rubber tube, sleeve, and metal fittings. In addition, the PAM isappropriate for applications that connect to the human body from the safety point of view because it can output
flexible and back-drivable driving force. The PAM contains the rubber tube and the outer sleeve knitted withfibers in an oblique direction. When air pressure is applied into the rubber tube, the PAM expands in the radial
direction and simultaneously contracts in the long axis direction. Along with this deformation, the contraction
force is generated in the long axis direction of the PAM. By controlling the air pressure applied to the PAM, theproposed device gives a variable load to the trainee's muscles. The length of the PAM is 800 mm. The innerdiameter of the PAM is φ9.5 mm and generates the contraction force of about 300 N when 400 kPa gaugepressure is applied to the PAMs of the initial length. In this paper, the pressure information is shown as gaugepressure.

Control of the Proposed Device
A pneumatic circuit for driving the proposed device is also shown in Fig. 2. The pneumatic circuit consists of acompressor, air tank, pressure regulator, servo valves (FESTO, MPYE5-1/4-010-B), pressure sensors (SMC,
PSE510-M5), a pipeline of air tubes, and PAMs. The pressure in the PAM is measured by the pressure sensor,
and its information (voltage) is sent to a control PC. The control voltage for the servo valve is calculated by thecomputer and is sent to the valve for controlling the pressure in the PAM. A block diagram of the pressure
feedback controller is shown in Fig. 3. The proposed training device gives the trainee variable load bycontrolling the pressure applied to the PAM. In Fig. 3, Pref [kPa] is the target pressure, and P [kPa] is themeasured pressure in the PAM. Kp [V/kPa] is the proportional gain, and it was set as 0.05 V/kPa experimentally.
Ki [V/(kPa∙s)] is the integral gain, and it was set as 0.05 V/(kPa∙s). u [V] is the control voltage, G [kg/s] is themass flow, and s is the Laplace operator.

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Figure 3. Block diagram of the pressure feedback controller.

In this study, we adopted a simple controller as shown in Fig. 3 because we focused on three main topics asshown in the section of contribution. However, the controller in Fig. 3 is primitive, and in addition, the internal
volume of the PAMs is large, so it is expected that the adopted pressure controller will cause a time delay of lessthan 1 s. In our previous study, we aimed to improve the control performance of power assist devices using
PAMs, modeled a pneumatic circuit system including PAMs, and improved the responsiveness of the pressure
control [15]. We consider the model in our previous study can also be implemented for the control system of the
proposed device. Quantitative evaluation of how the time delay amount due to the difference in control methods
affects the training results is an interesting research topic and we consider it to be one of our future works.

APPLICATION TO BALANCE BOARD TRAINING
Experimental Method
In order to confirm that the prototype device can be applied to actual training, we selected a balance board
training and carried it out as a verification experiment. In this training, the trainee is required to stand on an
unstable board and not to tilt the board largely. There will be various strategies not to tilt largely the balanceboard for each trainee. For example, there is a general strategy as follows. The trainee keeps its upper body
upright from the floor as much as possible so as not to tilt the board largely, and fine-tunes the position of thereaction force on the soles of both legs to maintain the body balance.
We customized our prototype device for the balance board training as shown in Fig. 4. A balance board andrigid frame with four push-button switches are set on the force plate. The trainee stands on the balance board,and force and torque on a contact point between the board and the force plate are measured as time-series data.
When the trainee tilts the board largely, the edge of the board pushes one or two buttons, and the PAMs are
actuated and support the trainee’s body until the board lifts off from the button. We expect that this control
strategy can improve the trainee’s body balance training method as follows. The proposed device can inform theboard's inclination state to the trainee by the PAM force, and it makes the trainee can recover to its stableposture quickly. By repeating this training, the trainee will be able to efficiently learn how to keep the bodybalance does not tilt the board.
We consider the proposed training device can be applied to train the muscular system and nervous system of any
part of the whole body by changing the fixing positions of the PAMs. Whether the muscular system or nervous
system is intensively trained depends on the exercise task to be performed. For example, in the case of balanceboard training considered in this study, the external force of the PAMs is used as a force presentation function
teaching the correct posture rather than a loading function. In this case, the physical load on the trainee is notlarge and the training effectiveness on the muscular system is small; however, this training is effective as anervous system training for improving the trunk balance.
Fig. 5 shows an interlocking relationship between the PAMs and switches. The board inclination direction is
distinguished by the combination of pushed switches, and the corresponding PAM is driven in each mode. In
Fig. 4 right, upper and lower PAMs in four directions are driven by a common servo valve and contract at thesame time. Assuming the upper and lower PAMs as one set as shown in the same colors in Fig. 4 right, theproposed device applies a horizontal force to the trainer's waist using four PAM sets (PAMs 1 to 4). For
example, when the board tilts toward the subject's toes and presses the CH 1 switch, the pressure is applied to
the PAMs 2 and 3 to raise up the subject's body. When the board presses the two buttons CH 1 and 2, the PAMs
3 is pressurized to reduce the tilt angle of the board. In the PAM control, we set the target pressure value of a
step function shape. When the button is not pushed, the PAM target pressure Pref is 0 kPa. While the button ispushed, the corresponding PAM contracts by setting the target pressure value Pref as 300 kPa.
We carried out experimental measurements of two methods, the training using the proposed device and theconventional training without the device, and compared these performances. The number of subjects was five ineach group. The experimental procedure is shown in Fig. 6. The experiment is divided into three stages: Pre Test,
Test, and Post Test. The trainee stands upright on the board and the line of sight is fixed to the front with theeyes open. In Pre Test and Post Test, holding the above state for 40 s without wearing the PAMs is defined as
one trial, and three trials are measured in each of Pre Test and Post Test. In addition, the experiments performed

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Figure 4. Experimental setup for the balance board training.

Figure 5. Interlocking relationship between the PAMs and switches.

Figure 6. Experimental procedure.

in the Test differ depending on the group. In the group of the proposed method, the external force of the eight
PAMs is applied to the waist of the trainee who holds the above state, and in the group of the conventional
method, the PAMs are not attached to the trainee. Under these conditions, 10 trials are measured for 40 s in eachtrial. We set a rest time over one minute between each trial. Time series data of three axes of force and torqueare measured in each trial using the force plate. In this experiment, we evaluate the amount of body shake by thestandard deviation of the time series data. When the body shake is large, the standard deviation is also large. On
the contrary, when the body shake is small, the measured value becomes close to constant and the standard
deviation also becomes small. We obtain the measured values of Pre Test and Post Test and compare thetraining effect in each group. Specifically, we calculate the following index ΔF [%] as:F  Fpos / Fpre  100.

(1)Where F pos is the average value of the standard deviations for three Post Test data, and Fpre is that for three
Pre Test data. This value is calculated for each of the three forces and torques in each trainee. Obtain the
average and standard deviation of ΔF from the data of five trainees in each group, and statistically compare thegroups using the Student's t-test. Furthermore, for 10 trials of Test data, we calculate the standard deviation foreach trial and calculates the slope of the standard deviation with respect to the number of trials. This slope iscompared between the groups, and it is clarified whether there is a change in the learning speed of the balancetask due to the difference in the training method.

Experimental Results
Fig. 7 shows the average and standard deviation of ΔF in each axis of force and torque. In Fig. 7 left, the bold
bar represents the average value and the thin bar represents the standard deviation. The vertical axis shows ΔF

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Figure 7. Average and standard deviation of ΔF in each axis of force and torque.

Figure 8. Sample measured Test data.

Figure 9. Slopes of the standard deviations of Test data between the groups.

and the horizontal axis shows groups of w and w/o PAMs. p means the p values calculated by the Student’s ttest. Fig. 7 right shows the coordinate system of forces and torques measured by the force plate. The origin of
the coordinate system was set on the center of the force plate's upper surface.
In addition, Fig. 8 shows the sample measured Test data. The standard deviations of forces and torques arecalculated in each trial, and the slope of the standard deviation is obtained by linear approximation. In this figure,the solid lines are the standard deviation of force calculated from the measured value. The dashed lines are thelinear approximation of the standard deviation, and we focused on these slopes which will mean learning curvesof the trainee. Fig. 9 shows the results of comparing the slopes of the standard deviations of Test data between
the groups. In this figure, the bold bar represents the average value and the thin bar represents the standard error.
The vertical axis shows the slope value and the horizontal axis represents each axis component of force andtorque in each group. The green and blue colors indicate the group of w and w/o PAMs. The symbols * and **indicate that the values are statistically significantly different at the significance level of 10 and 5% using thetwo-sided test.
We discuss the results of Fig. 7. For all data, ΔF was less than 100%, and the Post Test value was smaller than
the Pre Test value. From these results, it was shown that balance board training is effective for reducing the6

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The 11th JFPS International Symposium on Fluid Power 2020body shake amount. Regarding the comparison between groups, the values of the proposed method group (wPAMs) were smaller than those of the conventional method group (w/o PAMs) for the three-axis components offorces and torque Mx. Torques My and Mz were larger in the proposed method group than in the conventional
method group. As an overall tendency, the group of the proposed method tends to have a relatively small ΔF.
However, in the experimental results, no significant difference was obtained between the groups in any data.
The cause was assumed as that the number of samples (number of trainees) in each group was small of five, andthe individual differences in the results were large. As our future work, we consider measuring the larger
number of trainee's data and control of athletic ability in the groups.
Next, we discuss the results of Fig. 9. In all the results of force and torque, the slope of the standard deviation of
Test was larger in the group of the proposed method than that of the conventional method. In particular, the Fy
result showed statistically a significant difference at a significance level of 5%, and the Fz result showed asignificant difference at a significance level of 10%. From these results, we considered that the proposedtraining method may be able to obtain the practice effect more efficiently in a shorter time than the conventional
method.
As shown in Fig. 8, the standard deviations of force and torque do not always decrease linearly with the increasein the number of trials. This is a common tendency that is often seen in human experimental measurements, and
it is due to the large variation in the measured values for each subject and each trial. To solve this problem, it iseffective to set a sufficient number of trials and observe the macroscopic statistical tendency of the data. In thisstudy, we confirmed the significant differences between the groups in 10 trials. A clearer tendency will beobserved by increasing the number of trials. However, since an increase in the number of trials leads to anincrease in the experiment time and the large burden on the trainee, it is necessary to consider an appropriate
number of trials in the future.

CONCLUSION
In this paper, we proposed the novel whole body motion training device by multi-directional force input using
PAMs for total motor system enhancement. In addition, we developed the first prototype of the proposed device.
The prototype device was applied to the balance board training. The proposed device can inform the board's
inclination state to the trainee by the PAM force, and it makes the trainee can recover to its stable posture
quickly. By repeating this training, the trainee will be able to efficiently learn how to keep the body balancedoes not tilt the board. In the experimental validation, we measured floor reaction force and torque in two cases,(i) training with the PAM force, (ii) training without the PAM force. In this experiment, we evaluated theamount of body shake by the standard deviation of the time series data. The experiment was divided into threestages: Pre Test, Test, and Post Test. We obtained the measured values of Pre Test and Post Test and compared
the training effect in each case. Furthermore, for 10 trials of Test data, we calculated the standard deviation foreach trial and calculated the slope of the standard deviation with respect to the number of trials. This slope wascompared between the groups. As a result of the experiment, we confirmed that the proposed training will beeffective to improve the trainee’s core balance performance in a shorter time compared with the conventional
training.
As our future work, we will observe the change of the trainee's performance when the training using the
proposed device is performed over a long period about several months.

ACKNOWLEDGMENTS
This work was supported by the Bridgestone Corporation and KAKENHI (20K19594, 21H04544). This
research was also based on the Cooperative Research Project of the Research Center for Biomedical
Engineering. The authors would like to thank Shingo Oono, Ryo Sakurai, and Shintaro Yoshida of Bridgestone
Corporation for their support.

REFERENCES
1. GOODPASTER, B. H., et al., The Loss of Skeletal Muscle Strength, Mass, and Quality in Older Adults: The Health,
Aging and Body Composition Study, Journal of Gerontology: Series A, Vol. 61, No. 10, 2006, p. 1059-1064.
2. IANNUZZI-SUCICH, M., et al., Prevalence of Sarcopenia and Predictors of Skeletal Muscle Mass in Healthy, Older
Men and Women, Journal of Gerontology: Series A, Vol. 57, No. 12, 2002, p. 772-777.
3. JOZSI, A. C., et al., Changes in Power with Resistance Training in Older and Younger Men and Women, Journal ofGerontology: Series A, Vol. 54, No. 11, 1999, p. 591-596.
4. FRONTERA, W. R., et al., Strength Conditioning in Older Men: Skeletal Muscle Hypertrophy and Improved Function,
Journal of Applied Physiology, Vol. 64, No. 3, 1988, p. 1038-1044.

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5. HUNTER, G. R., et al., Resistance Training Increases Total Energy Expenditure and Free-Living Physical Activity inOlder Adults, Journal of Applied Physiology, Vol. 89, No. 3, 2000, p. 977-984.
6. ORR, R., et al., Power Training Improves Balance in Healthy Older Adults, Journal of Gerontology: Series A, Vol. 61,
No. 1, 2006, p. 78-85.
7. COOK, G., Movement: Functional Movement Systems. On Target Publications, 2012.
8. BALACHANDRAN, A. T., et al., Power Training Using Pneumatic Machines Vs. Plate-Loaded Machines to ImproveMuscle Power in Older Adults, Experimental Gerontology, Vol. 98, 2017, p. 134-142.
9. HEWETT, T. E., et al., The Effect of Neuromuscular Training on the Incidence of Knee Injury in Female Athletes, The
American Journal of Sports Medicine, Vol. 27, No. 6, 1999, p. 699-706.
10. CARIGNAN, C. R., et al., A Haptic Control Interface for a Motorized Exercise Machine, 2008 IEEE International
Conference on Robotics and Automation, 2008, p. 2055-2060.
11. https://jp.ddrobotec.com/en/dd-system-elite/ (accessed on 19 May, 2021)
12. https://www.technogym.com/skillrun.html (accessed on 19 May, 2021)
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14. NAKANISHI, T., et al., Development of an Arm Curl Machine with Variable Resistance Using Pneumatic Artificial
Rubber Muscle, 2019 IEEE/RSJ International Conference on Intelligent Robots and Systems, 2019, p. 3830-3835.
15.TAGAMI, T., et al., Pressure Control of a Pneumatic Artificial Muscle Including Pneumatic Circuit Model, IEEE Access,
2020, p. 60526-60538.

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Development of Robotic Forceps Driven by Soft Actuatorwith Built-In Displacement Sensor
Osamu Azami*, Takahiro KANNO*, Toshihiro KAWASE**,***, Maina SOGABE****,
Tetsuro MIYAZAKI**** and Kenji KAWASHIMA*****

Riverfield Inc., 20 Samon-cho, Shinjuku-ku, Tokyo, 160-0017 Japan(E-mail: osamuazami@gmail.com)**
Institute of Biomaterials and Bioengineering, Tokyo Medical and Dental University
2-3-10 Kanda Surugadai, Chiyoda-ku, Tokyo, 101-0062 Japan***Tokyo Institute of Technology, 4259, Nagatsuta-cho, Midori-ku, Yokohama 226-8503, Japan****Tokyo University, 7-3-1, Hongo, Bunkyo-ku, Tokyo 113-8656, Japan

Abstract. We have proposed a pneumatically driven extension-type soft actuator with a built-in displacement sensor.
The actuator is constructed by a silicone tube and a metal spring. The displacement is obtained by measuring thechange in spring inductance. Furthermore, the force acting on the actuator is estimated from the pressure and thedisplacement by utilizing the high backdrivability of the soft actuator. In this paper, we proposed and prototyped arobotic forceps using the soft actuator with a built-in displacement sensor. The robotic forceps has a gripper and 2
degrees of freedom joints at the tip. The joints angle and gripper of the robotic forceps are driven by pulling the wire.
The pushing force of the soft actuator is transmitting to the pulling force of the wire through the pulley. We evaluate
the characteristics of the forceps through experiments. There is a linear relationship between the forceps bending angle
and the actuator displacement. The acting force of the forceps tip can be roughly estimated from the supply pressureand displacement of the soft actuator.
Keywords: pneumatic, soft actuator, robotic forceps, built-in sensor

INTRODUCTION
Soft robots and actuators are more flexible and lighter than traditional rigid robots. Therefore, the robots andactuators are safe in contact with human [1]. Soft actuators are being actively studied [2] because the spread of
3D printers has made it easier to prototype. Various shapes of soft actuators have been proposed, and they areapplied in various fields such as industrial and medical fields. Regarding the driving method of soft actuators,
various methods are used, for example hydraulic, pneumatic and electric. In particular, pneumatic drive is suitable
for medical applications because it has the following features: power-to-weight ratio is high and working fluid isclean. Examples of pneumatically driven soft actuators include McKibben type artificial muscle and fiberreinforced actuators [3][4]. Fiber-reinforced actuator has a large stroke and the relationship between the generated
force and the supply pressure is generally linear. It is required sensors to operate the soft actuator precisely. Several
actuators with built-in sensors have already been proposed. However, general sensors have the following
problems: they are sensitive to heat or water and expensive. Therefore, we have proposed a new pneumatic soft
actuator with a built-in displacement sensor [5][6]. The actuator is constructed by a silicone tube and a metal
spring. The spring has two roles: reinforcing the outer periphery and sensing displacement. It is able to estimate
the actuator displacement by measuring the change in spring inductance. Furthermore, the actuator has high backdrivability since it is pneumatically driven. Therefore, the force acting on the actuator can be estimated from the
supply pressure and the displacement. We considered that the actuator is suitable for application in the medical
field since it has the following characteristic: right weight, low cost, and autoclavable.
In this paper, we proposed and prototyped a robotic forceps driven by the soft actuator. By using the proposed
actuator, a lightweight and simple structure forceps can be realized. We proposed a method to estimate the acting
force to the forceps tip from the pressure and displacement of the soft actuator. The static and force estimation
characteristics of the robotic forceps were examined experimentally.

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Figure 1. Prototyped robotic forceps driven by soft actuators with built in sensor.

Figure 2. Drive mechanism the robotic forceps.
(a) Crank mechanism. (b) Antagonistic drive.

ROBOTIC FORCEPS
Fig. 1 shows the prototyped robotic forceps. The total length is 485 mm, the drive unit length is 130 mm, and the
maximum outer diameter is 75.0 mm. The weight of the forceps is 196.5 g. Fig. 2 shows the schematic drawingof the drive mechanism. The driving force of the actuator is transmitted to the tip joint bending wire via the crankmechanism. As shown in Fig. 2(a), the crank mechanism is necessary to exert a pulling force on the bending wire,
since the soft actuator is an expansion type. The forceps mechanism operates by antagonistic drive, and as shownin Fig. 2(b), when an actuator extends, a corresponding actuator (on the opposite side of the same plane) must
contract in cooperation. Since the forceps are equipped with two sets of this mechanism, the forceps has twodegrees of freedom. The proposed actuator has the minimum length in its natural state. We have set a backlash at
the crank mechanism-actuator connection to achieve antagonistic drive. The pulley angle θp is 45 ° when theactuator displacement is 0, and the actuator driving force acts on the wire when θp is 60 ° or more. The backlash
converted to actuator displacement is about 2.3 mm. By geometrically calculating from the dimensions of eachpart, the acting force Ff of action point at the tip of the forceps can be expressed by Eq. (1).

Ff =

Lw
L r
Fw = w a FairLf
L f rw(1)

Here Lw is the length of the moment arm of the bending wire, Lf is the length from the base of the forceps tip joint
to the point of action, ra is radius of the actuator side pulley, rw is radius of the wire side pulley, Fw is the acting
force on the bending wire and Fair is the driving force of the actuator.

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Table 1. Dimensions of each part of the crank mechanism.
Length [mm]2.1
Lw

Dimension parameters

46.5

Lf
ra
rw

9.0
9.0

Figure 3. Pneumatically driven extension-type soft actuator(a) Soft actuator. (b) Soft actuator mounted on the forceps.

Table 1 shows the dimensions of each part of the mechanism. We set the maximum supply pressure to the actuator
to be 200 kPa. In our previous study, we have confirmed that the output force of the soft actuator was larger thanthe pure product of the inner cross-sectional area and the pressure since the rubber wall also transmitted pressureusing a thick-walled cylinder model [7]. In the pre-experiment, we measured the supply pressure P and generated
force F of the actuator, and calculated the effective cross-sectional area A by assuming Fair = P × A. As a result,
A is 2.409×10-4 m2. By calculating based on Eq. (1), the maximum value of Ff is determined to be 2.17 N.
Fig.3(a) shows the soft actuator with built-in displacement sensor. The actuator is constructed by a silicone tube
and a metal spring. The outer diameter of the actuator is 20mm and the natural length is 40mm. It extends 32.69mm
with a pressure of 200kPa. The spring has two roles; reinforcing the outer periphery and sensing displacement. It
is able to estimate the actuator displacement by measuring the change in spring inductance. Furthermore, theactuator has high back drivability since it is pneumatically driven. Therefore, it is able to estimate the force acting
on the actuator from the supply pressure and the displacement. We have experimentally confirmed that thedisplacement can be estimated with an error of 5.6%, and the external force can be estimated with an error of 13 %performing experiments with a low frequency sinusoidal input of 0.5 Hz or less [6]. The actuator mounted on theforceps is covered with a pipe with an inner diameter of 21 mm to prevent flexion movement. (Fig.3(b)).
We confirmed that the effects of autoclave sterilization (120 ° C, 30 minutes) were negligible. Therefore, the soft
actuator can be used to drive the forceps. It is possible to reduce the weight of the forceps, reduce the cost, andfacilitate sterilization by using the actuator for driving the robot forceps.

EXPERIMENTS
Static Characteristics of Tip Joint Angle
We conducted an experiment to measure the static characteristic of the actuator displacement x and the joint angle
q at the tip joint. Fig. 4 shows the experimental apparatus. The supply pressure to the actuator was controlledusing a servo valve (Valve box). The control signal to the valve was calculated with the PC. The displacement of
the actuator was measured from the inductance change of the spring. The joint angle was measured from the image
taken by a camera. The relationship between the actuator displacement xest and the joint angle q was derived fromthe experiments. The target value of the actuator displacement was set as follows: 0.0, 0.5, 1.0, 1.5, 2.0, 2.5, 3.0,
3.5, 4.0, 4.5, 5.0, 5.5 mm. Only one actuator was mounted on the forceps, the antagonistic drive was not implement.
The experimental results are shown in Fig. 5. From the results, it can be seen that when xest is less than 2 mm, q
remains in initial position, and when xest becomes larger than 2 mm, xest and q have linear characteristics.
Furthermore, we derived an approximate equation Eq. (2) to obtain q from xest.

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Figure 4. Apparatus for static characteristics experiments of robotic forceps.

Figure 5. Static characteristics between xest and q of developed robot forceps.

qest = 0.0( xest  2.0 )qest = 0.356xest − 0.681( xest  2.0 )(2)Force Estimation at the Forceps Tip
We conducted an experiment to estimate the acting force at the tip of the forceps from the state change in the soft
actuator. Fig. 6 shows the experimental apparatus. The actuator was driven by position feedback control, and theacting force Ff is measured by a force sensor. We calculated the estimated value Ffest of the acting force from theactuator driving force Fair using Eq. (1). The joint angle of contact with a force sensor was set as 0 and 10 °, andcompared Ff and Ffest.
The experimental results are shown in Fig. 7. Fig. 7(a) shows the result of Ff and Fig. 7(b) shows the result of Ffest.
The waveform of Ff and Ffest show a similar trend. However, the average error is 23% at a bending angle of 0 °,and 68% at a bending angle of 10 °. We consider that this error is mainly caused by the friction acting on a bending
wire inside the forceps shaft.

Figure 6. Experimental apparatus for measuring Ff.

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Figure 7. Results of forceps force.
(a) The force measured by a force sensor. (b) The estimated force from the pressure and displacement change of the actuator

CONCLUSION
In this paper, we proposed the robotic forceps driven by pneumatic soft actuator with a built-in sensor. The soft
actuator is lightweight, low-cost, and sterilizable by autoclave. Therefore, it is possible to reduce the weight of
the forceps, reduce the cost, and facilitate sterilization by using this actuator for driving the robot forceps. We
proposed a method to estimate the acting force to the forceps tip from the supply pressure and displacement of thesoft actuator. The static and force estimation characteristics of the robotic forceps were examined experimentally.
We confirmed that the relationship between estimated value of actuator displacement xest and forceps bending
angle q have linear characteristics. We confirmed that the trends of force acting on the tip for each condition are
the same between the measured value Ff and the estimated value Ffest. The average error is 23% at a bending angleof 0 °. However, the error is not small and increases at the bending angle of 10°. This may be due to the friction
between the outer cover and the soft actuator, the friction acting on the wire, and the unintended flexion movementof the actuator. The future work is to reduce the frictions, and increase the flexion rigidity of the soft actuator
during operation.

REFERENCES
1. D. Rus and M. Tolley: Nature 521 (2015) 467. https://doi.org/10.1038/nature14543
2. R.F. Shepherd. Multigait soft robot. PNAS. 2011, Volume 108, No. 51, 20400–20403..
3. E. W. Hawkes. Design and Implementation of a 300% Strain Soft Artificial Muscle. In 2016 IEEE International
Conference on Robotics and Automation, IEEE: Stockholm, Sweden, 2016; pp. 4022–4029.
4. P. Polygerinos. Modeling of Soft Fiber-Reinforced Bending Actuators. IEEE Transactions on Robotics 2015, Vol. 31,
Issue 3, 778-789.
5. O. Azami, Development of the extension type pneumatic soft actuator with built-in displacement sensor. Sensors &Actuators:A 2019, 300. 111623.https://doi.org/10.1016/j.sna.2019.111623.
6. O. Azami. External Force Estimation of Pneumatic Soft Actuator with Built-in Displacement Sensor. Sensors & Materials
2020, Volume 33, No. 2(1), pp. 555–567. https://doi.org/10.18494/SAM.2021.3152.
7. T. Kanno. Model of Coil-Reinforced Cylindrical Soft Actuator. Applied Sciences 2019, Volume 9, Issue 10, 2109

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Genaral Session | Oil hydraulics

Tribology, Seals, and Contamination Control[GS4-01] Effect of Sealing Surface Flatness on Leakage Characteristics of FlangeType Gasket Model Using Oil Viscosity-Temperature Relations〇Song Gao1, Toshiharu Kazama1 （1.Muroran Institute of Technology）[GS4-02] Experimental Analysis of Rotational Motion of Pistons and Slippers of a
Swashplate Axial Piston Pump Using Visualization Technique〇Takumi Furuya2, Toshiki Haga2, Toshiharu Kazama1 （1.Muroran Institute of Technology,
2.Graduate School of Muroran Institute of Technology）[GS4-03] Feasibility and Precision Analysis of a Test Rig with Adjustable Oil Film
Thickness〇Haiji Wang1, Guanglin Shi1 （1.Shanghai Jiaotong University）©The Japan Fluid Power System SocietyGS4-01

The 11th JFPS International Symposium on Fluid Power 2020

Effect of Sealing Surface Flatness on Leakage Characteristics ofFlange-Type Gasket Model Using Oil Viscosity-Temperature
RelationsSong GAO* and Toshiharu KAZAMA***Division of Engineering, Graduate School of Muroran Institute of Technology
27-1, Mizumoto-cho, Muroran, Hokkaido, 050-8585 Japan(E-mail: 19096005@mmm.muroran-it.ac.jp)**
College of Design and Manufacturing Technology, Muroran Institute of Technology
27-1, Mizumoto-cho, Muroran, Hokkaido, 050-8585 Japan

Abstract. Gaskets have a wide range of applications as sealing elements in industry, machinery, and daily life. They
prevent fluid leakage between stationary surfaces according to the connected components. Generally, leakage can be
restrained by pressing both sides of the sealing surfaces and blocking them up. However, reducing leakage is difficult
because roughness and waviness exist on the sealing surface, and deformation may occur due to strong clamping,which can cause partial channels. Therefore, maintaining a flat sealing surface during operation is difficult. In thisstudy, a theoretical model of flange-type gasket is developed using the thermo-hydrodynamic lubrication theory underflat, convex, and concave flange-surface conditions. The concept involves the use of the fluid characteristic that theviscosity increases as the temperature decreases. The effect of flange flatness on the temperature distribution, flow
velocity, and leakage was investigated for both conditions of parallel and inclined gap.
Keywords: Seal, Gasket, Temperature, Oil, Viscosity, Flatness

INTRODUCTION
In fluid power and mechanical engineering, sealing elements are vital and indispensable. Gaskets [1-2], astraditional and typical mechanical sealing elements, are widely applied in the industry and factory. Gaskets contain
artificial materials that include rubber, copper alloy, and plastic. These materials are installed between thecontacting surfaces to enhance the sealing performance of mechanical elements. Normally, leakage is positively
correlated to the gap space and is inversely proportional to the flange sealing area. However, excessive clamping
can lead to the damage of the sealing surface of gaskets, and elastic deformation may occur due to over extrusion.
According to the elastic deformation in theory of elasticity, Persson [3] proposed a theory for estimating a leakage,and sealing performance under both static and dynamic conditions, including plastic deformation, was developed.
Lehn et al. [4] analyzed a 3D thermo-elasto-hydrodynamic model of an air-foil thrust bearing by considering thethermal management, the temperature analysis and effect of deformation of the bearing surface on angular velocity
were discussed. Furthermore, realizing a flat sealing surface is difficult under practical applications. To achieve
outstanding sealing performance, strong clamping is necessary, which can result in surface deformation. Therefore,
in the design process of a gasket, we need to investigate the leakage from the gap using theoretical and qualitative
analyses of the parameters that affect leakages and apply methods to suppress them. Analysis of a gasket flangewith a non-planar sealing surface should also be performed. Gadari et al. [5] numerically and isothermallyanalyzed an elasto-hydrodynamic lubrication model of seals to estimate the effects of lip deformation assemblyand shaft roughness on a revolving lip seal. Shinkarenko et al. [6] developed a theoretical model to investigate
practical application of laser-surface texting in the soft elasto-hydrodynamic lubrication. They stated that rigidcounterpart surface testing could increase the load-carrying capacity and reduce friction, which led to an optimal
capacity in terms of the aspect ratio and area density with preferred dimples.
In this study, we focus on the lubricant viscosity, particularly that of oil, which is strongly related to temperature.
Low temperature corresponds to high fluid viscosity [7]. The fluid viscosity can be increased by cooling the gasket
sealing flange. Then, leakage can be reduced. Gao and Kazama [8] proposed the thermo-hydrodynamic lubrication(THL) and iso-viscous methods to control leakage using the characteristics of a flange-type gasket. The flatnessof a flange surface is also an important factor that affects the sealing performance. In this paper, the phenomenon
between the leakage and solid wall temperature, and the relationship between cooling of sealing flange and leakage
reduction is theoretically verified. The THL theory is applied to analyze gap flow, temperature, flow velocity, andleakage, and the effects of temperature control, gap space, and pressure on leakage are calculated using the finite
difference method (FDM). The theoretical solution focuses on several types of sealing-surface shapes, e.g., flat,
convex, and concave surfaces. The effect of surface flatness on the temperature distribution in the lubricating film,
flow velocity, and leakage under both parallel and inclined gap conditions are primarily discussed.

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THEORY
Figure 1 shows the schematic diagram of a flange-type gasket model with two circular flanges. 𝑃𝑟 is the central
recess pressure, 𝑃𝑒 is ambient pressure, 𝑇𝑤 is the temperature of the solid wall, 𝑇𝑖𝑛 is oil temperature, 𝑇′𝑟 is central
temperature of disk, 𝑇𝑎 is ambient temperature. Flange thickness 𝐵 is fixed at 𝑟 = 𝑅1 . The THL theory is applied
to the gap flow with considering the sealing-surface flatness under static conditions of parallel and inclination.

Figure 1. Schematic Diagram of the Flange-Type Gasket Model.

The coupled differential equations contain the generalized Reynolds equation [Eq. (1)], energy equation of the oilfilm [Eq. (2)], heat conduction equation of upper flange [Eq. (3)], and equations of the lubricant physical properties.
1 ∂
𝑟 ∂𝑟

𝜌𝑐𝑝 (𝑢

(𝐹2 𝑟

∂𝑇
∂𝑟

1 ∂
𝑟′ ∂𝑟′

+

∂𝑝
∂𝑟

𝑣 𝜕𝑇
𝑟 𝜕𝜃

(𝑟′

∂𝑇′
∂𝑟′

1 ∂

)+ 2

𝑟 ∂𝜃

+𝑤

)+

∂𝑇
∂𝑧

(𝐹2

1 ∂2 𝑇′∂𝜃∂)=𝑟′2 ∂𝜃 2

∂𝑝

∂𝑧

+

)=0

(𝜆

∂𝑇

) + 𝜇𝜙

(2)

=0

(3)

∂𝑧

∂2 𝑇′
∂𝑧′2

(1)

where 𝑐𝑝 is the specific heat capacity of lubricating oil and 𝜆 is the thermal conductivity. Functions of 𝐹0 , 𝐹1 , 𝐹2 ,dissipation function 𝜙, and boundary conditions of pressure and temperature coincide with those in Reference [8].
Gap space ℎ, flow velocity 𝑢, and leakage 𝑄𝑜𝑢𝑡 from the gap at the edge of disk (𝑟 = 𝑅2 ) are expressed as follows:ℎ = ℎ𝑐 + 𝛼𝑟cos(𝜃 − 𝜑) + 𝛿 (
𝑢=

∂𝑝

𝑟−𝑅1
𝑅2 −𝑅1

𝑧𝑧

𝐹

𝑧 d𝑧

𝜇

𝐹0

𝜇

(∫0 d𝑧 − 1 ∫0

)

2

(4)

)

(5)

𝑄𝑜𝑢𝑡 = ∫0 ∫0 𝑢𝑅2 d𝑧 d𝜃

(6)

∂𝑟

2π

ℎ

where 𝛼 is the inclination angle of the gap space and 𝛿 is the surface-flatness coefficient. 𝛿 < 0 indicates that theflange surface is concave. 𝛿 = 0 indicates a flat surface. 𝛿 > 0 indicates the convex flange surface. The viscosity–
temperature relationship in the lubricating oil is expressed as2

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𝜈 = 𝜈0 exp[−𝛽(𝑇 − 𝑇0 )](7)

where 𝛽 is the oil viscosity–temperature index, which is a constant associated with temperature 𝑇0 . In this study,
viscosity is a single function of temperature, and the other oil physical properties, including density, coefficientsof specific heat, and thermal conductivity, can be considered constant. The Reynolds, energy, and heat conduction
equations, are coupling solved using FDM, and the equations are calculated by iteration.
The upper and lower disks are classified as A5052 in JIS and SUS303, respectively. The selected disk parameters
are as follows: inner and outer radii are 𝑅1 = 60 and 𝑅2 = 100 mm, and disk thickness 𝐵 = 10 mm at position𝑟 = 𝑅1 . The lubricant property is selected as that of ISO VG32 oil using calculation. The other input parameters
are as follows: 𝜌 = 869 kg⁄m3 , 𝜈0 = 32.6 × 10−6 m2 ⁄s , 𝛽 = 0.02969 [1⁄℃] at 𝑇0 = 40 ℃ , 𝑐𝑝 = 1860
J⁄(kg ∙ K) , 𝜆 = 0.130 W⁄(K ∙ m) , 𝜇 = 𝜌𝜈, 𝜆′ = 16.3 W⁄(K ∙ m) , ℎ𝑡 = 50 W⁄(m2 ∙ K) , and 𝑇𝑖𝑛 = 𝑇′𝑟 = 𝑇𝑎 =20 ℃. The effects of the surface shape on gasket sealing performance such as the temperature distribution, flow
velocity, and leakage are presented in the Results and Discussion section.

RESULTS AND DISCUSSION
Theoretical Solutions of the Effect of 𝜹 on the Temperature Distribution
Figure 2 shows that the gap space ℎ changes with three kinds of flatness coefficient 𝛿 along radius under parallel
gap with a central gap of ℎ𝑐 = 0.15 mm. We observe that clearance increases with a larger value of 𝛿 along the 𝑟direction. In this study, all theoretical calculations are performed under the conditions of ℎ𝑐 = 0.15 mm and 𝑃𝑟 =1.4 kPa. The effect of 𝛿 on the gasket sealing performance with circular flanges is discussed relative to these three
types of sealing surface as shown in Fig. 2.

Figure 2. Variation in Gap Space ℎ with Radius 𝑟 under a Parallel Gap Condition (ℎ𝑐 = 0.15 mm).

Temperature Solutions under a Parallel Gap Condition
Figures 3–5 show the theoretical results of the temperature contour map in the section of upper flange and oil film
for parallel flange with flat (𝛿 = 0), convex (𝛿 = 0.05), and concave (𝛿 = −0.05) sealing surfaces, respectively.
The upper and lower parts indicate the flange and oil domains. The temperature in the abscissa (𝑧 = 0) represents
the solid wall temperature (𝑇𝑤 = 5 ℃). Although the temperature values at each grid node in the lubricant film
are slightly different, they can be feasibly controlled by 𝑇𝑤 in all three cases.

Figure 3. Theoretical Result of 𝑇 and 𝑇′ with a Flat Flange Derived by the THL Theory (𝛿 = 0).

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Figure 4. Theoretical Result of 𝑇 and 𝑇′ with a Convex Flange Derived by the THL Theory (𝛿 = 0.05).

Figure 5. Theoretical Result of 𝑇 and 𝑇′ with a Concave Flange Derived by the THL Theory (𝛿 = −0.05).

Figure 3 shows the temperature distribution for a flat flange (𝛿 = 0). Figure 4 shows the convex flange condition(𝛿 = 0.05). The gap space increases to ℎ = 0.2 mm at the edge of disk at 𝑟 = 𝑅2 . It decreases to ℎ = 0.1 mm in
the concave flange (𝛿 = −0.05), as shown in Fig. 5. We need to mention that in order to intuitively observe thetemperature contour in the cross section of flange and oil film, disk thickness 𝐵 and gap space ℎ are plotted at
equidistant positions in Figs. 3–5. However, the gap space changes along the radial direction at 𝛿 ≠ 0. Therefore,
we show the specific values of the gap space at the oil inlet (𝑟 = 𝑅1 ) and outlet (𝑟 = 𝑅2 ) in Figs. 3–5.
Figures. 3–5 show the contour line of 𝑇 = 7.50 ℃ and 7.92 ℃, which is marked in red, to indicate the positionsof the same contour in three cases. By comparing the result with the flat flange case, the locations of the marked
contours point downward in the convex flange (𝛿 = 0.05), and the overall temperature in the oil film is relatively
deviated from target value 𝑇𝑤 . These values shift upward in the concave flange case (𝛿 = −0.05), and the oil film
temperature approaches 𝑇𝑤 . In other words, the temperature of oil film area can be easily controlled by the solid
wall using a flange with a concave sealing surface.
Temperature Solutions under the Condition of Gap with an Inclination Angle
Figures 6–8 show the temperature contour in the interface between the gap and the upper flange, i.e., 𝑇𝑧=ℎ (𝑇′𝑧 ′ =0 )under an inclined flange condition (𝛼 = 0.057°) with flat (𝛿 = 0), convex (𝛿 = 0.05), and concave (𝛿 = −0.05)
sealing surfaces, respectively. The solid wall temperature is set at 𝑇𝑤 = 5 ℃. Although the gap–disk interface is
not a flat circle because of the axial deformation at 𝛿 ≠ 0, as shown in Figs. 7 and 8, they are still expressed in acircular form for comparison with the flat case shown in Fig. 6.
We can observe that the small variations in temperature exist in the gap–disk interface, which is affected by thedifferent values of flatness coefficient 𝛿. The temperature at the interface agrees with that of the solid wall, andthe temperature point near the minimum gap position (ℎ𝑚𝑖𝑛 ) the most coincides with 𝑇𝑤 in all the three cases. In
comparison with the flat flange, the overall temperature at the interface is generally deviated from the solid wall
temperature 𝑇𝑤 in the gap with the convex flange. However, it is closer to 𝑇𝑤 in concave case, as shown in Fig. 8.
The overall temperature values in the oil film was much more theoretically agrees with the target value 𝑇𝑤 in thesmaller gap space condition.

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Figure 6. Temperature Contour in the Gap–Disk Interface in an Inclined Gap with a Flat Flange (𝛿 = 0).

Figure 7. Temperature Contour in the Gap–Disk Interface in an Inclined Gap with a Convex Flange (𝛿 = 0.05).

Figure 8. Temperature Contour in the Gap–Disk Interface in an Inclined Gap with a Concave Flange (𝛿 = −0.05).

Theoretical Solutions of the Flow Velocity and Leakage
Flow Velocity under a Parallel Gap Condition
Figures 9–11 show the distribution of flow velocity 𝑢 along the radius in terms of the 𝑧 value under parallel gap
using flange with flat (𝛿 = 0), convex (𝛿 = 0.05), and concave (𝛿 = −0.05) sealing surfaces, respectively. The
solid wall temperature is 𝑇𝑤 = 5 ℃. The specific values of the gap space at the oil inlet and outlet are also shown.
Compared with the flat flange case as shown in Fig. 9, the flow velocity of with convex sealing surface decreases
near the oil outlet, as shown in Fig. 10, and the closer the surface is to the outlet, the larger is the decrease in thegradient. Meanwhile, the flow velocity of with concave sealing surface increases, as shown in Fig. 11. The velocity
difference in three cases can be mainly attributed to the difference in the rate of change in gap along 𝑟 direction.

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Figure 9. Distribution of Velocity 𝑢 for a Flat Sealing Surface (𝛿 = 0).

Figure 10. Distribution of Velocity 𝑢 for a Convex Sealing Surface (𝛿 = 0.05).

Figure 11. Distribution of Velocity 𝑢 for a Concave Sealing Surface (𝛿 = −0.05).

Leakage Flow under Both Conditions of Parallel and Inclined Gap
Figure 12 shows the effects of flatness coefficient 𝛿 of the flange surface on the leakage flow under parallel gap
condition. Figure 13 shows the effects of 𝛿 on the leakage flow under the inclined gap with 𝛼 = 0.057°. Solidwall temperature 𝑇𝑤 is varied from 5 ℃ to 45 ℃, and the temperature interval is 10 ℃.
Comparison of the parallel and inclined gap cases reveals that the leakage increases under the inclined case at thesame values of ℎ𝑐 and 𝑃𝑟 . The leakage also increases with larger flatness coefficient 𝛿, and the leakage–restraint
phenomenon is most obvious in the flange with concave surface in both cases of parallel and inclination, as shownin Figs. 12 and 13. The main reason is that the a larger flatness coefficient corresponds to a larger gradient in gap
space along the 𝑟 direction, i.e., the gap space is larger at the edge of disk (𝑟 = 𝑅2 ) while the center gap remains
the same. Another reason is that when the gap is larger, controlling the oil film temperature by solid wall is moredifficult. However, although the leakage increases with a larger flatness coefficient, it still can be suppressed bylowering the temperature in the flanges with different types of sealing surfaces under parallel and inclined gaps.

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Figure 12. Effect of 𝛿 on the Leakage Flow under Parallel Gap Conditions (𝛼 = 0°).

Figure 13. Effect of 𝛿 on the Leakage under Inclined Gap Conditions (𝛼 = 0.057°).

CONCLUSIONSIn this study, a theoretical model for analyzing the temperature, flow velocity, and leakage of a flange-type gasket
was developed using the viscosity–temperature relationship. The THL theory was applied to the gap flow under
the sealing surface of installed flat, convex, and concave flanges. The leakage and oil film temperature weretheoretically calculated using the surface flatness coefficient and inclination angle were considered as the main
parameters. When the temperature of the sealing flange was controlled, the leakage varied under the range of theflatness coefficient and inclination angle. The conclusions of this study are presented as follows:(1) The oil film temperature could theoretically agree with that of solid wall under various conditions independentof the shape of sealing surface and parallel configuration of the gap.
(2) Comparing to flat flange, the flow velocity in the concave flange theoretically increased along radius for bothconditions of parallel and inclined gap. The closer the outlet was, the larger was the gradient increase.
(3) Irrespective of the inclination angle, the leakage could be most obviously restrained by lowering temperatureof the sealing parts by using a flange with a concave surface.

REFERENCES
1.
2.
3.
4.

5.
6.
7.

8.

JIS B0116, Glossary of Terms of Packing and Gaskets, 2015.
ASTM F118, Standard Definitions of Terms Relating to Gaskets, American Society for Testing and Materials, 1985.
PERSSON, B. N., Leakage of Metallic Seals: Role of Plastic Deformations, Tribology Letters, Vol.63, 2016, 42.
LEHN, A., MAHNER, M., SCHWEIZER, B., A Thermo-Elasto-Hydrodynamic Model for Air Foil Thrust Bearings
Including Self-Induced Convective Cooling of the Rotor Disk and Thermal Runaway, Tribology International, Vol.119,2018, pp.281-298.
GADARI, M. E., FATU, A., HAJJAM, M., Shaft Roughness Effect on Elasto-Hydrodynamic Lubrication of Rotary Lip
Seals: Experimentation and Numerical Simulation, Tribology International, Vol.88, 2015, pp.218-227.
SHINKARENKO, A., KLIGERMAN, Y., ETSION, I., The Effect of Surface Texting in Soft Elasto-Hydrodynamic
Lubrication, Tribology International, Vol.42, 2009, pp.284-292.
KAZAMA, T., GAO, S., Gap Flow Between Two Circular Plates with Temperature-Controlled Wall: Application of
Thermohydrodynamic Lubrication Theory and Comparison with an Iso-Viscous Model, Mechanical Engineering Letters,
Vol.6, 2020, 20-00292.
GAO, S., KAZAMA, T., Leakage Control for a Flat Flange Model with a Gap Based on Sealing Liquid Viscosity, Journalof Advanced Mechanical Design, Systems, and Manufacturing, Vol.15, No.3, 2021, 21-00065.

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Experimental Analysis of Rotational Motion of Pistons and Slippers
of a Swashplate Axial Piston Pump Using Visualization Technique
Takumi FURUYA*, Toshiki HAGA*, and Toshiharu KAZAMA***Division of Production Systems Engineering, Graduate School of Muroran Institute of Technology
27-1 Mizumoto-cho, Muroran, Hokkaido, 050-8585 Japan**
College of Design and Manufacturing Technology, Graduate School of Muroran Institute of Technology
27-1 Mizumoto-cho, Muroran, Hokkaido, 050-8585 Japan(E-mail: kazama@mmm.muroran-it.ac.jp)
Abstract. The spinning motion of the pistons and slippers of a swashplate-type axial piston pump was investigatedusing a visualization technique. The pump casing was partially replaced with transparent acrylic resin plates. The
pistons and slippers with the cylinder block and the swashplate were visualized using a high-speed video camera.
Images were extracted from the video and were analyzed, and their behavior was also examined experimentally. The
shaft rotational speed and discharge pressure were as high as 1500 rpm and 5 MPa, respectively. By changing thepump’s mounting posture, the motion of the pistons and slippers were observed from almost all circumferential
directions to cover both discharge and suction ports. The experimental results revealed that the changes in the slipper
spin exceeded those in the piston spin.
Keywords: Hydraulic pump, Piston assembly, Spin, High-speed camera, Tribology

INTRODUCTION
A swashplate-type piston pump can handle high pressures while ensuring a high degree of safety, compactness,and efficiency [1]. The piston assembly of the pump comprises pistons and slippers, which are connected with aball joint. The piston reciprocates in a cylinder bore with the swashplate [2], and the slipper is attached to theother end of the piston to support high and fluctuating loads while sliding at low to high speed on theswashplate’s sliding surface [3].
The pistons and slippers also rotate around the main driving shaft while spinning around the cylinder-bores’ axis.
The rotation and spin behaviors of the pistons and slippers influence the pump’s reliability and performance.
Few studies have investigated this subject theoretically and experimentally [4]. However, the motion of theseparts remains sufficiently unclarified.
Herein, the spinning motion of the piston and slipper is visualized using a high-speed video camera, and their
behavior is experimentally examined.

EXPERIMENT
Test Apparatus and Method
Figure 1 shows the hydraulic circuit of the test apparatus [5], which comprises a test pump, a three-phaseinduction motor (7.5 kW), an electric inverter, a torque transducer (20 Nꞏm), a flow meter, pressure sensors,thermometers, an oil cooler, hydraulic valves, and a reservoir.
The test pump was developed by modifying the commercial swashplate axial piston pump of a rotating cylinderblock type. The basic specifications of the original pump were as follows: maximum pressure = 21 MPa,maximum rotational speed = 1800 rpm (30 s−1), theoretical discharge = 10 ml/rev, and number of pistons = 9.
Also, transparent acrylic windows were installed in the revamped casing to observe the motion of the pistonsand slippers. The piston assemblies were numbered from one to nine to differentiate between each piston andslipper. Moreover, each piston and slipper were marked with a permanent marker pen to track the rotating
motion.
Mineral oil-type hydraulic fluid (density: 869 kg/m3 at 15°C; kinematic viscosity: 32.6 mm2/s at 40°C and 5.49mm2/s at 100°C) was used as the test oil.

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M
O

T

P

A

T

B

M

Figure 1. Schematic representation of apparatus’ hydraulic circuit.

Image Analysis and Test Conditions
A high-speed video camera and two metal halide lamps were used to record the motion. The distance between
the video camera’s lens and the test pump’s window was set as 310 mm, and the lamps were placed at both sidesof the camera. The frame rate was 4000 fps, and the shutter speed was 1/20000 s.
The oil temperature was maintained at 30°C. The pump’s rotational speed and discharge pressure were set up to1500 rpm (25 s−1) and 5 MPa, respectively. The motion of the pistons and slippers was observed and recorded at
the center of the discharge (0 rad) and suction ( rad) ports and at postures of −/4, /4, 3/4, and 5/4 rad.
Because a reference location was required to analyze the motion, a jig of a metallic stick was used before
recording the motion (Fig. 2). Different images were extracted from the video, which enabled reading theposition coordinates (Fig. 3). The images were extracted for every 1 and 10 frames of the conditions at 150 and
1500 rpm, respectively. To avoid errors based on the objects’ curvature, the data when the pistons and slippers
were crossing at the centerline of the main shaft were adopted. Thus, the number of frames for the image
analysis was 12–18 each.

Figure 2. Reference position indexing in images.

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Figure 3. Measuring points of pistons and slippers extracted from a video frame (unloaded, 300 rpm (5 s−1), full swashplateangle, and 0°-pump posture).

RESULTS AND DISCUSSION
Figures 4 and 5 show the rotational motion of the pistons and slippers under the conditions of full and halfswashplate angles, respectively. The pump’s rotational speed was set as 2.5 and 25 s−1, and the dischargepressure was set at unloaded (slightly > atmospheric pressure) and 5 MPa. The parameter is the ratio of thespinning speed of the pistons and slippers to the rotational speed of the pump’s main shaft. Thus, a zero value ofthe ratio indicates that the spinning speed coincides with the shaft speed, and a positive value implies that the
spinning speed exceeds the shaft speed and vice versa. In Figs. 4 and 5, the left (pump angle from −/2 to /2)
and right (pump angle from /2 to 3/2) halves approximately correspond to the discharge and suction ports,
respectively.
For all conditions, the fluctuation of the ratio of the slippers exceeded that of the pistons. Comparing Figs. 4 and5, the fluctuation of the full angle (Fig. 4) was much marked than that of half-angle (Fig. 5). Besides, the ratios
and their fluctuations became higher at the centers of both the discharge and suction ports, especially for theslippers under full swashplate angle condition.
Moreover, the slippers’ spinning motion of the unloaded condition and that of the 5 MPa condition almost
coincided in the half-angle case (Fig. 5, right-hand side), implying that the discharge pressure less affected theslippers’ spinning. Overall, for conditions of full and half swashplate angles, the mean values of the ratios of theslippers were positive, whereas the values of the pistons were almost zero.

a)b)

Figure 4. Spin ratio comparison of pistons and slippers for maximum swashplate angle (a: piston; b: slipper).

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Figure 5. Spin ratio comparison of pistons and slippers for half swashplate angle (a: piston; b: slipper).

CONCLUSIONS
The motion of the pistons and slippers was measured using a visualization technique. Under the present
conditions, the variation of the slippers’ rotation exceeded that of the pistons, and the rotational speed of theslippers was slightly higher than that of the shaft over the entire circumference.

ACKNOWLEDGMENTS
The authors are grateful to Yuken Kogyo Co., Ltd. for supplying the test pump.

REFERENCES
1. YAMAGUCHI, A., TANAKA, H., Hydraulics and Pneumatics, 1986, Corona Pub. Co., Ltd. (in Japanese)
2. KAZAMA, T., Comparison of Temperature Measurements and Thermal Characteristics of Hydraulic Piston,
Vane, and Gear Pumps, Mechanical Engineering Journal, 2015, 2-3, pp.14-00542.
3. KAZAMA, T., SUZUKI, M., SUZUKI, K., Relation between Sliding-Part Temperature and Clearance Shapeof A Slipper in Swashplate Axial Piston Motors, JFPS International Journal of Fluid Power System, 2015, 81, pp.10-17.
4. CHAO, Q., WANG, Q., XU, B., CHEN, Y., LI, Y., Experimental Investigations of the Slipper Spin in an
Axial Piston Pump, Measurement, 2017, 102, pp.112-120.
5. HAGA, T., KAZAMA, T., Visualization Measurement of Rotational Motion of Pistons and Slippers of A
Piston Pump (Preliminary Experiment), Proceedings of MoViC2019, 2019, C102. (in Japanese)4

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Feasibility and Precision Analysis of a Test Rig with Adjustable Oil
Film Thickness
Haiji WANG* and Guanglin SHI **School of Mechanical Engineering, Shanghai Jiaotong University
800 Dongchuan Rd, Shanghai, 200240 Japan(E-mail: glshi@263.net)

Abstract. In the present study, a test rig with adjustable oil film thickness for the slipper is proposed. The
mathematical model of the oil film thickness, pressure and temperature is demonstrated. The mechanism of the test
rig including the test system, hydraulic system, control system and data collection is shown. The feasibility of thetest rig shows that the oil film thickness of the slipper can be precisely adjusted. It’s found that the relative errors of
the reacting force, torque and temperature in the experiment and simulation are mainly due to the deformation of theretainer. The secondary cause is the pressure fluctuation and temperature increasing from the input oil. The relative
errors of the characteristics are less when the simulation is modified. Based on the test rig, further works such as
testing the optimized slipper and finding the minimum energy consumption of different operating conditions areprepared to fulfill.
Keywords: Feasibility and precision, Adjustable oil film thickness, Slipper, Test rig.

INTRODUCTION
Slipper/swash-plate pair is the most significant lubricating interface in the piston pump. A suitable oil film
forms to separate heavily loaded relatively movable parts in each other[1]. Many researches and optimizationsof the piston pump mentioned the importance of the lubrication characteristics[2-5]. Therefore, a test rig fordetecting the lubrication characteristics is valuable.
Before 21st Century, Hooke and Kakoullis fulfilled a test rig to measure the oil film thickness with thecapacitance displacement sensor[6]. They analyzed the influence of centrifugal force and friction of ball joint
between piston and slipper on the lubrication of slipper/swash-plate pair. Based on this test rig, Koc measured
the oil film thickness of slipper/swash-plate pair with different sizes of orifice, eccentricity and overclamp
ratio[7-10].
After 21st Century, more researchers carried out excellent works on lubrication of slipper/swash-plate pair. Ai
designed a test rig which is different from Prof. Hooke’s test rig[11-13]. They can continuously measure thethickness since the displacement sensor placed on the retainer of the slipper. Torque loss of the slipper can bemeasured based on Canbulut’s test rig in order to evaluate the oil film thickness[14,15]. Manring placed threepressure sensors on the swash-plate, in order to describe the pressure distribution of the slipper/swash-platepair[16]. In order to carry out the oil film thickness and torque resulting from the tilting slipper, Rokala used
three displacement sensors and force sensor in the test rig[17,18]. This kind of test rig is improved by Xu[19-23].
Displacement sensor, torque sensor and pressure sensor are all placed to find out the slipper motion in axial
piston pump. Schenk finished a test rig with six displacement sensors o the swash-plate in order to get the
instantaneous oil film thickness[24].
From now on, there are two methods to measure the oil film thickness. The displacement sensor can be placed in
the swash-plate or it can be placed in the retainer of the slipper. Assume that the measurement rig is a system,
the size of the slipper and the operating condition are the input of this system, and the lubrication characteristics
such as oil film thickness, pressure distribution, torque even temperature are the output. However, oil film
thickness has great effect on the other lubrication characteristics. Therefore, the oil film thickness should be theinput of measurement rig. In the above test rigs, the oil film thickness can be measured, but it can’t be controlled.
In this paper, authors propose an adjustable oil film thickness test rig. The oil film thickness can be adjusted bythree couples of step motor and universal joint coupling. The accuracy of the test rig is determined by a threecoordinate measuring machine and three displacement sensors. The simulation of lubrication characteristics isfinished in Matlab. The results from test rig are compared with the results from the simulation. The reasons forthe errors between the experiment and simulation are analyzed. It’s found that the main cause of the errors is thedeformation of the retainer. After the modification of the simulation, the relative errors between the experimentand modified simulation reduce. Compared with authors’ last research[25], the temperature effect on the test rig
is considered in this paper.

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DESCRIPTION OF OIL FILM
Oil Film Thickness
The thickness distribution of oil film of slipper/swash-plate pair is the basic characteristics. The effect ofoverturning moment makes a tilt when the slipper slides on the swash-plate. Therefore, the thickness of eachpoint of the slipper sealing land is different. Oil film thickness is determined by the height of three points.
Figure 1 shows the location and height of three points. Point A, point B and point C are located on the outermost
side of the slipper sealing land. The angle between these points is 120 . The heights of each point
are h1 , h2 and h3 separately. Since the coordinates of three points are (0, R, h1 ),(3
1
3
1
R,- R, h2 ),( R,- R, h3 ) , Eq.
2222(1) shows the equation of oil film.

x

y

z 1

0

R

h1 1

3
1
R - R h2 1 = 0
2
2
3
1
R - R h3 1
2
2

(1)

The oil film thickness distribution in Cartesian coordinate and Polar coordinate is described in Eq. (2) and Eq.
(3).

1
1
1
h( x, y )  3 ( h2  h3 ) x (2 h1  h2  h3 ) y  ( h1  h2  h3 )
R
3R
3
1
1
1
h( r , )  3 ( h2  h3 )r cos (2 h1  h2  h3 )r sin   ( h1  h2  h3 )
R
3R
3

(2)

(3)

Figure 1. Oil film thickness model 1.

In Figure 2, the oil film thickness distribution can also be described by the center height of the slipper hc , the tilt
angle  and the phase angle  . Eq. (4) shows the thickness distribution by this approach. The thickness andshape are clearly demonstrated. Referred to Eq. (1) and Eq. (4), the relationship between the height of threepoints and the center height of the slipper, the tilt angle and the phase angle is shown in Eq. (5). The thickness
and shape of oil film can be determined by Eq. (4). The height of three points can be adjusted by three step
motors of test rig. Therefore, it’s realized that the thickness and shape of oil film can be determined by the test2

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rig. The lubrication characteristics in different conditions are measured by the test rig.

h  hc  R tan  cos(   )(4)πh1  hc  R tan  cos(   )2πh2  hc  R tan  cos(    )67π
h3  hc  R tan  cos(   )6(5)

Figure 2. Oil film thickness model 2.

Oil Film Pressure
Researchers studied a lot on oil film pressure distribution. An incompressible Reynolds equation models the
distribution of pressure under the slipper sealing land. Eq. (6) shows the Reynolds equation in Polar coordinate.
The hydrostatic effect is illustrated on the left hand side of Eq. (6). On the right hand side, the first two parts
express the hydrodynamic effect meanwhile the third part expresses the squeeze effect.

1   3 p  1   3 p 
h
h 
 h
 vsθ
2 
 rh

h
  6   vsrr r 
r  r 2    

r


t 


(6)

The lubrication characteristics such as reacting force and torque are determined by the pressure distribution. Eq.
(7), and Eq. (8) show the lubrication characteristics separately[26].

F  pr πr02  

2π0Tor  2π

0R0 h p vsr R0

 prdrd(7)

r0

 h p vsθ 

 R  2 r  h  cos   2r   h  sin    rdrd
r0(8)

Eq. (9) shows the power loss. The first part on the right hand of Eq. (9) is the power loss of the leakage. Thep psecond part is the power loss of the torque. Tor is the integral of pressure gradient， and viscosity
r 
v v
friction sr ， sθ from r0 to R0 and 0 to 2π .
hh

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Oil Film Temperature
The oil film is regarded as a control volume[27]. The heat generation results from the power loss[23]. The
conductive heat transfer is occurred between the oil film and solid parts such as the slipper and swash-plate. The
radiative heat transfer is occurred between the solid parts and ambient air. Assume that the oil is incompressible,
Eq. (10) expresses the dynamic temperature of oil film[28].


dT
1 

  cp q Tin  T   Q  W 
dt  cpV (10)

Figure 3 demonstrates the temperature status between the oil film, slipper and swash-plate. According to
Newton’s law of cooling, the relationship of the temperature between the oil film, slipper, swash-plate andambient air is expressed in Eq. (11) and Eq. (12) separately.

T  T12 T11  T12 T12  T0


Rc1
Rc2
Rc3
T  T21 211  T22 T22  T0


Rc4
Rc5
Rc6
In Eq. (12), Rc1   k0 2πr0 H1 1

, Rc2  ln( R0 / r0 ) /  k1 2πr0 H1  , Rc3   k0 2πR0 H1 

(11)
(12)
1

. In Eq. (13), Rc4   k0 As1 

1,

Rc5  H 2  k2 As1  , Rc6   k0 As2  . Therefore, the heat transfer ratios of between the oil film and ambient air through1

1

the slipper and swash-plate are expressed in Eq. (13) and Eq. (14). The rate of heat rejection between the oil film
and ambient air is expressed in Eq. (15). Therefore, the entire rate of heat rejection of the control volume isexpressed in Eq. (16).

T  T0Rc1  Rc2  Rc3
T  T0Q2 
Rc4  Rc5  Rc6
Q1 

(13)
(14)

Q3  k0 2πR0 hc T  T0 

(15)Q  Q1  Q2  Q3

(16)

Figure 3. Oil film temperature model.

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INTRODUCTION OF TEST RIG
Mechanism of Test Rig
Figure 4 and Figure 5 show the main mechanism of the test rig. Servo motor on the left hand side controls therotating rate of swash-plate. Two couplings and torque sensor are placed between them. Tapered rolling bearing
and thrust rolling bearing are utilized in the end case. They can resist the torque and force resulting from theslipper to the swash-plate separately.
On the right hand side, three step motors control the movement of the retainer. Three couples of bar, universal
joint coupling and force sensor are placed between them. The slipper in blue connected with the piston is
clamped in the center of the retainer by a circlip. Three eddy current displacement sensors in yellow are
distributed over the circumference of the slipper. The laser temperature sensor in red can measure the surface
temperature of the slipper.
The schematic of the test rig is built as shown in Figure 6. It mainly includes hydraulic system, cooling system,
test system, control system and data logging system. The black curve shows the oil path of the test rig.
Meanwhile the red curve and blue curve show the control system and data collection. The test system is driven
by a servo motor. The hydraulic system including the servo motor, gear pump, valves and so on offers highpressure oil to the test system. The relief valve and throttle valve adjust the pressure and flow rate of the inlet
flow. The unloaded valve protects the hydraulic system when the pressure is too high. All signals obtained fromthe sensors are collected by a group of ADVANTECH DAQ modules. A Labview program is utilized to control
the system and record the data. Besides, the status of test rig is shown in the monitor. Figure 7 shows the test rig
in real.

Figure 4. Configuration of test rig.

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Figure 5. Cross section of key components.

Figure 6. Schematic of measurement.

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Figure 7. Test rig and components a) shows the entire test rig, b) shows the key components on the left hand side,c) shows the key components on the right hand side, d) shows the hydraulic system.

Table 1 lists the key components of the test rig and their main index. The key components of the hydraulicsystem are listed on the left hand side. On the right hand side, the range and accuracy of sensors are listed.

Description
Servo Motor M1
Servo Motor M3
Gear Pump
Relief Valve
Throttle Valve
Unloaded Valve
Torque Sensor
Pressure Sensor
Temperature Sensor
Force Sensor
Displacement Sensor
Step Motor

Table 1. Parameters of objective axial piston pump.
Index
2000 r/min
3000 r/min
0.66 mL/r
Max 20 MPa
Max 1 L/min
Max 25 MPa
Range
0-0.5 Nm
0-30 MPa

Accuracy
0.0015 Nm
0.03 MPa0  200 C
0-500 N
0-5 mm
0.3 mm

2C
0.5 N
1 m
1.5 m

Feasibility of Test Rig
The slipper is clasped by the retainer. Meanwhile the retainer, universal joint coupling and step motor can be
regarded as a spatial mechanism in 3D coordinate system. The feasibility of motion is decided by the input ofsystem (hereinafter referred to as ‘Input’) and the degree of freedom (hereinafter referred to as ‘DOF’). Grubler
analyzed the relationship between Input and DOF[29]. The mechanism doesn’t work when Input is more than
DOF; The position of motion can’t be guarantee when Input is less than DOF; When Input is equal to DOF, themovement and rotation can be precisely controlled.

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Figure 8 shows a brief spatial mechanism of test rig. According to Figure 4 and Figure 5, the number of Input is
3. Seven parts are placed in this mechanism, which are three couples of L-shaped slider and step motor, three
universal join coupling and a retainer with bars. L-shaped slider forms sliding pair with ground. Universal joint
coupling forms spherical pair with L-shaped slider and retainer separately. Table 2 lists DOF and constraint of
this spatial moving mechanism.
gEq. (17) shows the DOF of spatial mechanism.  fi is the sum of DOF of each pair. For example, DOF of eachi 1spherical pair is 3 and there are 6 spherical pairs in this spatial mechanism. Therefore, the entire DOF ofspherical pair is 18. Based on Eq. (17), DOF of the proposed spatial mechanism is 3, which is equal to thenumber of Input. Refer to Figure 4, the movement of retainer along the X axis and the rotation along the Y axis
and Z axis can be precisely controlled by three step motors.
g

M  6  ( N  g  1)   f ii 1

9 6  (7  9  1)   f i

(17)

i 1

3

Figure 8. Spatial mechanism.

MEASUREMENT RESULTS
Experiment Introduction
Figure 9 presents the process of simulation and experiment. The simulation is fulfilled in MATLAB. The
process of experiment is more complex. It has four steps. In the first step, the thickness between the slipper andswash-plate (oil film) is adjusted by three step motors. After that, run the servo motor of test rig. Record theresults from torque sensor and displacement sensor. In this time, the torque is caused by bearing. Therefore, thetorque resulting from oil film is the difference between the torque in step 4 and torque in step 2. In the third step,
run the servo motor of hydraulic system. The pressurized oil flows to the test system. There’s the reacting force
from the slipper sealing land. Since there are very small gaps in universal joint couplings, the shape andthickness of oil film should be adjusted again slightly by the step motors. Finally, all the results such as force,
torque and so on can be achieved.

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Figure 9. Process of simulation and experiment a) shows the process of simulation, b) shows the process of experiment.

Results Comparison
The experiment results and simulation results are compared. Temperature, viscosity, reacting force and torqueare shown in Figure 10. The oil film thickness is set to 5 μm . The operating pressure and rotating rate are set to
5 MPa and 1000 r/min separately. The running time of the test is 10 minutes. The oil film temperature can’tmeasure directly. However the slipper outer surface temperature T12 can be measured by the laser temperature
sensor. Based on Eq. (12), the oil film temperature can be evaluated.
In Figure 10, the red curves show the results from the simulation meanwhile the blue curves show the resultsfrom the experiment. According to Figure 10, the tendency of the evaluated oil film temperature is similar withthe simulated one, but the increasing rate is lower. The maximum relative error is about 13.4% when theexperiment is end. The accounts of the reacting force from the experiment and simulation are close. The
maximum relative error is about 1.2%, which is quite small. Temperature has no effect on the reacting force.
Higher temperature reduces the viscosity. Therefore, the torque from the experiment and simulation bothdecreases. The maximum relative error of torque is about 10% when the experiment is beginning. With timegoes on, the decreases rate from the experiment is lower than the one from the simulation. The reason is that thetemperature increasing rate from the experiment is lower than the one from the simulation.
The error of the reacting force is small. However other errors are larger than 10%. The reasons are from many
aspects. In the next section, the reasons for the errors are analyzed. The simulation model is modified in order tomeet the true status of the test rig.

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Figure 10. Results comparison between the experiment and simulation a) shows the oil film temperature,b) shows the oil film viscosity, c) shows the reacting force, d) shows the torque.

ERROR DISCUSSION
Test Rig Assembly Error
All parts of the test rig are turned, drilled and milled. With the development of the manufacturing technology,
the accuracy and quality of parts are getting higher and higher. However, the errors in assembly are unavoidable.
They have great impact on the accuracy of the test rig. Referred to Figure 4, the errors in assembly along the Yaxis and Z axis should be focused. No matter where the errors occur in assembly, they eventually lead toconcentricity errors of two end cases. Therefore, relative position between swash-plate and slipper in test rig ischanged. Figure 11 shows the ideal and wrong relative position between swash-plate and slipper. The error
between ideal position and wrong position causes wrong radial shear velocity and wrong circumferential shear
velocity. The pressure distribution and reacting force are wrong as well. Therefore, the effect of the errors along
the Y axis and Z axis should be calculated.
Based on the results from three-coordinate measuring machine. The maximum errors along the Y axis and Z axis
are 0.0654 mm and 0.0143 mm separately. Therefore, the maximum eccentricity error is 0.0669 mm, which
influences the pitch radium . Table 2 lists the maximum relative errors between the experiment and different
simulations.
Based on Table 2, the difference of the maximum temperature relative error between the experiment andmodified simulation is larger. However, other relative errors are almost same. According to the results, the test
rig assembly error is not the main cause of the characteristics errors.
Table 2. Maximum relative errors between the experiment and different simulation 1.

Description
Temperature
Reacting force
TorqueWithout Eccentricity
11%
1.7%
10%

With Eccentricity
13.6%
1.7%
10.1%

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Figure 11. Ideal position and wrong position between the slipper and swash-plate.

Operating Condition Error
In the simulation, the pressure and temperature of the input oil are constant. In the experiment, the operatingpressure is generated by the hydraulic system. Although the gear pump is utilized since it can generate low
fluctuating pressurized oil, the fluctuation of operating pressure leads to the errors of the characteristics. Besides,
the temperature of the input oil from the gear pump increases linearly. Therefore, import the experimentalpressure and temperature data into the simulation. The results are shown in Table 3.
Based on Table 3, the maximum relative error of the reacting force and temperature is larger since the input
pressure data for the modified simulation is overshoot sometimes. Therefore, operating condition error is thesecondary cause of the characteristics errors.
Table 3. Maximum relative errors between the experiment and different simulation 2.

Description
Temperature
Reacting force
TorqueConstant Operating Condition
11%
1.7%
10%

Real Operating Condition
13.6%
5.1%
10%

Retainer Deformation
The retainer withstands the force from the slipper. It deforms the center of retainer along the X axis, which leads
to the tested displacement error. In order to reduce the effect of the deformation, multiple thickness retainers are
simulated in ANSYS. The maximum deformation is about 8 μm when the thickness of the retainer is 3 mm. It
reduces to 0.9 μm when the thickness is 6 mm, which is smaller than the accuracy of the eddy current
displacement sensor. Considering the deformation of the retainer in the simulation, the characteristics are shownin Table 4.
Based on Table 4, the main reason of the difference between the experiment and simulation is the deformation of
the retainer. The oil film thickness of the experiment and simulation is set to 5 μm . However, because of thedeformation, the oil film thickness raises to about 5.9 μm . Therefore, when it’ set to 5.9 μm in the simulation.
The errors of the characteristics reduce.
Table 4. Maximum relative errors between the experiment and different simulation 3.

Description
Temperature
Reacting force
TorqueWith Deformation
11%
1.7%
10%

Without Deformation
5.8%
1.7%
6.7%

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CONCLUSIONS
A novel oil film thickness adjustable test rig is proposed. The mathematical models of the oil film thickness,pressure and temperature are illustrated. It’s characteristics such as the reacting force, torque, and temperaturefrom the experiment and simulation are compared.
The mechanism of the test rig including the hydraulic system, cooling system, control system and data logging
system are introduced. The feasibility of the mechanism is calculated that the thickness and shape of the oil film
can be precisely controlled by the mechanism.
Based on the initial results from the experiment and simulation, the error of the reacting force is less than 10%.
However, the errors of other characteristics are larger. The simulation is modified in order to match the realsituation of the experiment.
The assembly error of the test rig is taken into account in the simulation. However, it’s not the main cause of thecharacteristics errors. The main cause is the retainer deformation and the secondary cause is the pressure
fluctuation and input oil temperature increasing. The characteristics errors reduce when the retainer deformation
and real operating condition are considered in the simulation.

FURTHER WORK
A novel oil film thickness adjustable test rig is made for testing oil film characteristics in the different operating
conditions. The optimized slipper is designed in the simulation and it can be tested in this test rig. After that,based on the results from the test rig, an oil film thickness adjustable valve-control axial piston pump is
designed which can work with minimum energy consumption in the different operating conditions.

NOMENCLATURE
As1
As2
cp
F
g
H1
H2
h13
hc
k0
k1
k2
M
N
n
pr
q

: Swash-plate inner area
: Swash-plate outer area
: Heat capacity: Reacting force: DOF of each pair
: Slipper thickness: Swash-plate thickness: Oil film thickness in each point: Central oil film thickness: Coefficient of oil heat transfers: Coefficient of slipper heat transfers: Coefficient of swash-plate heat transfers: Degree of freedom: Number of parts: Rotating rate
: Pocket pressure: Leakage



Q
R
R0
Rc1-c6r0
T0: Rate of heat rejection: Pitch radius: Slipper outer radius: Thermal resistance: Slipper inner radius: External oil temperature

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T11
T12
T21
T22
V
vsr
vsθ

: Slipper inner surface temperature: Slipper outer surface temperature: Swash-plate inner surface temperature: Swash-plate outer surface temperature: Control volume: Radial velocity: Axial velocity



W







: Power loss
: Tilt angle: Phase angle: Rotating angle: Oil density: Oil viscosity

REFERENCES[1] CHAO, Q., ZHANG, J., XU, B., WANG Q., Discussion on the Reynolds Equation for the Slipper Bearing
Modeling in Axial Piston Pumps, Tribology International, Vol. 118, 2018, pp. 140-147.
[2] CHO, I. S., A Study on the Optimum Design for the Valve Plate of a Swash Plate-Type Oil HydraulicPiston Pump, Journal of Mechanical Science and Technology, Vol. 29, No. 6, 2015, pp. 2409–2413.
[3] HONG, Y. S., LEE, S. Y., KIM, S. H., LIM, S. Y., Improvement of the Low-Speed Friction Characteristics
of a Hydraulic Piston Pump by PVD-Coating of Tin, Journal of Mechanical Science and Technology, Vol.
20, No. 3, 2006, pp. 358–366.
[4] KIM, J. K., JUNG, J. Y., Measurement of Fluid Film Thickness on The Valve Plate in Oil Hydraulic Axial
Piston Pumps (I)-Bearing Pad Effects, Journal of Mechanical Science and Technology, Vol. 17, No. 2, 2004,
pp. 246–253.
[5] KIM, J. K., KIM, H. E., LEE, Y. B., JUNG, D. Y., SH, O. H., Measurement of Fluid Film Thickness on the
Valve Plate in Oil Hydraulic Axial Piston Pumps (Part I Spherical Design Effects), Journal of Mechanical
Science and Technology, Vol. 19, No. 2, 2008, pp. 655–663.
[6] HOOKE, C. J., KAKOULLIS, Y. P., The Effects of Centrifugal Load And Ball Friction on the Lubricationof Slipper in Axial Piston Pumps, 6th International Fluid Power Symposium, Cambridge, England, 1981,
pp. 179-191.
[7] KOC, E., LI, K. Y., HOOKE, C. J., Slipper Balance in Axial Piston Pumps and Motors, Transactions of the
ASME, Vol. 114, No. 10, 1996, pp. 766-772.
[8] KOC, E., HOOKE, C. J., Investigation into The Effects of Orifice Size Offset and Overclamp Ration on
the Lubrication of Slipper Bearings, Tribology International, Vol. 29, No. 4, 1996, pp. 299-305.
[9] KOC, E., HOOKE, C. J., Considerations in The Design of Partially Hydrostatic Slipper Bearings,
Tribology International, Vol. 30, No. 11, 1997, pp. 815-823.
[10] KOC, E., HOOKE, C. J., An Analysis of the Lubrication Mechanisms of the Bush-Type Bearings in High
Pressure Pumps, Tribology International, Vol. 30, No. 8, 1997, pp. 553-560.
[11] AI, Q., ZHOU, H., ZHANG, Z., YANG, H., Development of Lubricating Characteristics Testing System ofFriction Pair in Axial Piston Pump, Chinese Hydraulics & Pneumatics, Vol.11, 2004, pp. 22-25.
[12] ZHUANG, Q., ZHOU, J., AI. Q., Development of Experiment Rig for Lubricate Film of the Slipper-Pad in
Axial Piston Pumps, Machine Tool & Hydraulics, Vol. 3, 2005, pp. 113-114.
[13] ZHUANG, Q., ZHOU, J., AI. Q., Comparison in Measurements of Lubricating Film of the Slipper-Pair in
Axial Piston Pump, Machine Tool & Hydraulics, Vol. 12, 2005, pp. 86-88.
[14] CANBULUT, F., The Experimental Analyses of the Effects of the Geometric and Working Parameters on
the Circular Hydrostatic Thrust Bearing, JSME International Journal Series C, Vol. 48, 2005, pp. 715-722.
[15] CANBULUT, F., SINANOGLU, C., KOC, E., Experimental Analysis of Frictional Power Loss ofHydrostatic Slipper Bearings, Industrial Lubrication and Tribology, Vol. 61, No. 3, 2006, pp. 123-131.
[16] CRABTREE, A., MANRING, D., JOHNSON. R., Pressure Measurements for Translating Hydrostatic
Thrust Bearing, International Journal of Fluid Power, Vol. 6, No. 3, 2005, pp. 19-24.

13

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The 11th JFPS International Symposium on Fluid Power 2020[17] ROKALA, M., CALONIUS, O., KOSKINEN, K., Study of Lubrication Conditions in Slipper-Swash Plate
Contact in Water Hydraulic Axial Piston Pump, Proceeding of the 7th JFPS International Symposium on
Fluid Power Toyama, Japan, 2008.
[18] ROKALA, M., CALONIUS, O., KOSKINEN, K., Tribological Conditions between Swashplate andSlipper Pad in Variable Displacement Water Hydraulic Axial Piston Unit, Proceeding of the 6thInternational Fluid Power Conference Dresden, Germany, 2008.
[19] LI, Y., XU, B., Axial Piston Pump Slipper Pad Dynamic Characteristics of Wedge Oil Film, Chinese
Hydraulics & Pneumatics, Vol. 9, 2010, pp. 87-91.
[20] LI, Y, Study on Characteristics of Oil Film for Slipper Pair of Axial Piston Pumps, Master Thesis, Zhejiang
University, Hangzhou, China, 2001.
[21] XU, B., LI, Y., ZHANG, B., ZHANG. J., Numerical Simulation of Overturning Phenomenon of Axial
Piston Pump Slipper Pair, Journal of Mechanical Engineering, Vol. 46, No. 20, 2010, pp. 161-168.
[22] XU, B., ZHANG, J., YANG, H., Investigation on Structural Optimization of Anti-Overturning slipper of
Axial Piston Pump, Science China Technological Sciences, Vol. 55, No. 11, 2012, pp. 3010-3018.
[23] XU, B., CHAO, Q., ZHANG, J., LI. Y., Slipper Optimization Model Based on Equilibrium Coefficient,
Journal of Zhejiang University, Vol. 49, No. 6, 2015, pp. 1009-1014.
[24] SCHENK, A., Predicting Lubrication Performance between the Slipper and Swashplate in Axial Piston
Hydraulic Machines, Ph.D Thesis, Purdue University, Indiana, USA, 2014.
[25] WANG, H. J., SHI, G. L., An Adjustable Oil Film Thickness Test Rig for Detecting Lubrication
Characteristics of Slipper/Swash-Pair in Piston Pumps, Industrial Lubrication and Tribology, Vol. 73, No.1,
pp. 190-198.
[26] WANG, Y. J., Study on Performance of Slipper Bearing in Hydraulic Axial Piston Pump in High-Pressureand High-Speed, Ph.D Thesis, Beijing Institute of Technology, Beijing, China, 2014.
[27] SHI, G. L., WANG, H. J., Thermal-Hydraulic Model FOR Axial Piston Pump with Leakage and Friction
and its Application, Industrial Lubrication and Tribology, Vol. 71, No. 6, 2019, pp. 810-818.
[28] BUSQUETS, E., IVANTYSYNOVA, M., Temperature Prediction of Displacement Contrilled MultiActuator Machines, International Journal of Fluid Power, Vol. 14, No. 1, 2013, pp. 25-36.
[29] GRUBLER, M., Allgemeine Eigenschaft en der Zwanglaufigen Ebenen Kinematische Kette, Civiling
Enieus, Vol. 29, 1883, pp. 167-200.

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Genaral Session | Pneumatics

Components and Systems 1[GS5-1-01] Magnetic Sensor Study for Improving Air Turbine Spindle Performance〇Vanisara Kaewnamchai1, Tomonori Kato1, Kazuki Kawakubo1, Kazumasa Yamashita1（1.Fukuoka Institute of Technology）[GS5-1-02] Study on Multi-Cylinder Type Wind Powered Air Compressor
Applied Hypocycloid〇Ryota Tanoue1, Toshinori Fujita1 （1.Tokyo Denki University）[GS5-1-03] Development of Bidirectional Arm Curl Machine Using Pneumatic
Artificial Rubber Muscles〇Toshihiro Kawase1,2, Tomoya Nakanishi1, Shintaro Yoshida3, Shingo Ohno3, Ryo Sakurai3,
Tetsuro Miyazaki4, Takahiro Kanno5, Maina Sogabe4, Yoshikazu Nakajima1, Kenji Kawashima4（1.Tokyo Medical and Dental University, 2.Tokyo Institute of Technology, 3.Bridgestone
Corporation, 4.The University of Tokyo, 5.RIVERFIELD, Inc.）[GS5-1-04] Design of a Pneumatic Oscillator for Paper Machine's Doctor Blade
Systems
Stefano Colaiuda1, 〇Michele Gabrio Antonelli2, Pierluigi Beomonte Zobel2, Massimiliano
Centofanti1 （1.SMC Italia S.p.A., 2.Department of Industrial and Information Engineering andEconomics, University of L'Aquila）©The Japan Fluid Power System SocietyGS5-1-01

The 11th JFPS International Symposium on Fluid Power 2020

Magnetic Sensor Study for Improving Air Turbine Spindle
Performance
Vanisara KAEWNAMCHAI *, Tomonori KATO*,
Kazuki KAWAKUBO*, Kazumasa YAMASHITA**

Graduate School of Engineering,
Fukuoka Institute of Technology
3-30-1 Wajirohigashi, Higashi-ku, Fukuoka, 811-0295, Japan(E-mail: ad20102@bene.fit.ac.jp)

Abstract. Air turbine spindle control systems, in conjunction with high-precision quick-response pneumatic pressure
regulators, have been developed for various purposes, such as tool wear investigations and medical machine
applications. However, some limitations arose in the speed sensor (tachometer) used in the air turbine system. The
tachometer cannot function properly in low-speed conditions (less than 1,000 min-1). Therefore, this study aimed to
improve the low-rotational-speed measurement system. Moreover, a magneto-rheological fluid (MRF) damper was
proposed for preliminary experiments to enhance the air turbine, which will be included in a future reinstallation of
the entire system. Two commercial Hall effect sensors are installed to measure the rotational speed of the air turbine
spindle shaft and detect the magnetic field density in the MRF damper. The results showed that the Hall effectsensors could be used to achieve the proposed intention.
Keywords: Pneumatics, Air turbine spindle, Hall effect sensor, Functional fluid, MRF damper.

INTRODUCTION
In the past decade, the rotation feedback control system known as the high-precision quick response pneumaticpressure regulator (HPQR) [1] has been used in several applications in conjunction with the air turbine spindle
due to its high speed, low vibration, and low friction. Therefore, the air turbine spindle and HPQR system isessential for ultra-high-speed machines, such as milling machines and medical devices. [2-4]In previous studies [5-6], the authors improved the air turbine spindle to investigate tool wear by milling sample
plates. The system showed satisfactory performance when the rotational speed was high. However, thetachometer, a non-contact sensor placed on the air turbine spindle to measure the rotational speed, had anunacceptably significant error for a rotational speed of less than 1,000 min-1.
To solve this problem, we propose a Hall effect sensor for measuring the low rotational speeds of an air turbine
spindle. Furthermore, this study focuses on the preliminary experiments of magneto-rheological fluid (MRF)
dampers to enhance the air turbine spindle by using a Hall effect sensor to detect oil viscosity inside the damper
at various magnetic field densities. Some studies have investigated the viscosity of oil lubricants using Hall
effect sensors. For instance, Sriratana and Murayama [7] used a Hall effect sensor to measure lubricant viscosity
by varying the weight of metal particles. Similarly, Satthamsakul et al. [8] studied a Hall effect sensor that
detected oil lubricant deterioration under variations in the electromagnetic field generated by a ferrite-coresolenoid coil. In addition, Li et al. [9] present a rotary MR damper design in which a magnetic field was used togenerate electrical and structural parameters that increased the rotor working area and reduced the moment ofinertia of the rotor. The electromagnetic field of the MR damper was analyzed using the finite element method
in a simulation of the torque-current properties of the MR damper.
In addition, numerous studies have used Hall effect sensors for inspecting non-destructive testing, such as
Sriratana and Murayama [10], who investigated the imperfection of a material by comparing different
placements of the Hall effect sensor module. Furthermore, other researchers have studied motor control andhave detected relevant parameters using Hall effect sensors. For example, Fernandez et al. [11] discuss thecontrol of a permanent magnet synchronous motor (PMSM) using a Hall effect sensor to measure the position.
The device commonly used to measure the rotor position in a PMSM is a position sensor or an encoder/resolver,
however, these have drawbacks in terms of the costs and space required for installation. In conclusion, Hall
effect sensors are broadly used in various applications because of their small size, low cost, high sensitivity,
wide temperature range, and high performance.

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PRINCIPLESHall Effect Sensor
Hall effect sensors rely on the Hall effect theory for applications in various sensing device objectives. The
Lorentz force (F) is exerted on an electrically charged particle (q) that is moving with a certain velocity (v)through an electric field (E) and a magnetic flux density (B). The Lorentz force is defined as shown in Eq. (1)𝐹 = 𝑞[𝐸 + (𝑣 × 𝐵)](1)

The Hall voltage (VHall) is generated by the external magnetic field (B) when a constant current (I) passes
through the semiconductor plate (Hall element). The thickness of the Hall element is defined by d, and RHall isthe Hall coefficient, as shown in Eq. (2). [12]𝑉𝐻𝑎𝑙𝑙 =

𝑅𝐻𝑎𝑙𝑙
𝐼𝐵
𝑑

(2)

In this study, digital and analog Hall effect sensors were applied to the system to overcome the limitations of the
air turbine spindle rotational speed control. The commercial Hall vane sensor consists of a permanent magnet
and a digital Hall effect sensor fabricated in a sealed plastic package.
The Hall vane sensor provided a digital signal output. Typically, when the magnetic field is not present, theoutput signal is in an OFF state, as this Hall vane sensor package contains an internal permanent magnet
mounted in a twin tower. Therefore, the output signal from this Hall vane sensor is in the ON state when there isno ferrous gear interrupted.
The analog Hall effect sensor provides an analog output voltage proportional to the magnetic field density, when
no magnetic field is present, known as a null voltage. When the Hall effect sensor is sensed by the South Pole,the output voltage increases above the null voltage, while the output voltage decreases below the null voltage
when the North Pole is sensed.

MRF Damper
The magneto-rheological fluid (MRF) can be changed from a normal liquid to a semi-solid state by applying
magnetic fields. The rotor with the cup shape is connected to the rotating shaft of the damper, placed between
the MRF and the coil. When the coil inside the damper is excited, a magnetic field occurs around the damper.
Subsequently, the metal particles in the MRF performed as the magnetic field undergoes the gap between thedamper surface and the rotor. The metal particles generated a force restricting the rotating shaft of the damper.
Therefore, the rotating torque can be controlled by generating a current that is applied to the coil. Figure 1 (a)and (b) show the MRF behavior when no magnetic field exists and when the coil generates a magnetic field,
respectively.
Rotor

Magnetic field path

Coil

MRF
MRF damper
surface(a)(b)

Figure 1. MRF damper behavior.
(a) Magnetic field is not generated.
(b) Magnetic field is generated.

TESTING SYSTEM INSTALLATION
The two different Hall effect sensors are placed in different parts. First, a Hall vane sensor is installed with an
air turbine spindle shaft to measure the low-speed control condition. Second, an analog Hall effect sensor isinstalled with the rotary MRF damper for detecting the magnetic fields.
The diagram of the expected system installation is shown in Figure 2. This diagram will be followed when
installing the combination of MRF damper devices after testing this preliminary experiment.

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Rotary MRF damper
Tachometer

Analog Hall effect sensor

Air supply for air
turbine spindle

Air turbine spindle

Hall vane sensor
Air supply foraerostatic bearing
Figure 2. System installation.

Hall Vane Sensor Installed in the Air Turbine Spindle

Output voltage of Hall vane sensor V [V]The installation of the Hall vane sensor with the gear mounted on the air turbine spindle is shown in Figure 3.
Figure 4 shows the output voltage square waveform of the Hall vane sensor for a target speed of 1,000 min-1,
0.0601 second per 10 pluses related to the ten teeth of the gear that interrupted the magnetic field density in one
cycle. The frequency of the rotation speed is 16.63 Hz, so the rotation speed from the Hall vane sensor isapproximately 998 min-1. Therefore, the resolution of the Hall vane sensor is ± 2 min-1. In addition, theoperating parameters of the Hall vane sensor with the air turbine spindle are listed in Table 1.
Air turbine spindle

Gear

Hall vane sensor

12
10
8
6
4
2
0
0.00

0.02

0.04

0.06

Time t [s]Figure 4. Output voltage waveform of Hall vane sensor
for a target speed of 1,000 min-1.

Figure 3. Installation of the Hall vane sensor andthe target gear on the air turbine spindle.

Table 1. Operating parameters of the Hall vane sensor with the air turbine spindle.
Parameters
Rotation speed target

Value
500 min-1

Operating voltage
Operating current (Maximum at 20 ºC)
Operating temperature range

12 VDC
10 mA-20 to 85 ºC

Analog Hall Effect Sensor Equipped with MRF Damper
The analog Hall effect sensor was installed on the side surface of the MRF damper, where the coil was woundinside. Figure 5 shows the testing installation of the MRF damper, along with the analog Hall effect sensor’s
shape and size compared to those of a 1 Yen coin.
The operating voltage of this analog Hall effect sensor was 5 VDC, and the output voltage of Hall effect sensor
when there is no magnetic field was approximately 50% of the operating voltage (i.e., null voltage). The
operating parameters of the analog Hall effect sensor with the MRF damper are shown in Table 2.

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Analog Hall effect sensor

MRF damper

Figure 5. Installation of the analog Hall effect sensor with the MRF damper.
Table 2. Operating parameters of the analog Hall effect sensor.
Parameters

Value

Operating supply voltage

5 VDC

Sensitivity (at 25 ºC)

3.125 mV ± 0.125 mV/gauss

Operating temperature range
Null (Output @ 0 Gauss, V)-40 to 150 ºC
2.50 ± 0.075 V

EXPERIMENTAL RESULTS
The results were divided into two main parts and one additional part. We evaluated the performance between thedigital Hall vane sensor and tachometer and the analog Hall effect sensor to clarify the possibility of using acommercial MRF damper with an air turbine spindle system.
The additional capacitive displacement sensor experiment was conducted to measure the amplitude while testing
the Hall vane sensor and tachometer when the air turbine spindle is rotating. For the additional experiment, thecapacitive displacement sensor was installed along the shaft of the air turbine spindle, leaving a small gap
between sensor and shaft.
Hall vane sensor
Tachometer

Amplitude A [mm]

4

2

0

-2

-4
0

1

2

3

4

5

6

Time t [s]Figure 6. Amplitude of the tachometer and Hall vane sensor.

As shown in Figure 6, the amplitude of the Hall vane sensor is higher than that of the tachometer, which isapproximately 1.25 μm. It is found that the Hall vane sensor has little impact on the air turbine spindle shaft due
to the magnetic force from the permanent magnet that is placed in the twin tower of the Hall vane sensor.
In the experimental results of the magnetic fields detected by the analog Hall effect sensor, the factors discussed
are the output voltage of the analog Hall effect sensor and the supply voltage of the MRF damper to excite themagnetic fields. The experiment started from zero voltage applied to the MRF damper to voltages of 1, 2, 3, …,
22, 23, and 24 VDC. As shown in Figure 7, the output of the Hall effect sensor increases when increases thevoltage supply of the damper. Although no voltage to the MRF damper is supplied, the Hall effect sensor output
voltage does not change from the null voltage. Note that the results show a linear trend, which would relate to
the viscosity of the MRF.

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Output voltage of Hall effect sensor [V]The 11th JFPS International Symposium on Fluid Power 2020

2.54
2.53
2.52
2.51
2.50
2.49
2.48
2.47
2.46
0

2

4

6

8 10 12 14 16 18 20 22 24 26

Voltage supply of MRF damper [V]Figure 7. Relationship between the analog Hall effect sensor and MRF damper.

The experimental results in terms of the step response for rotational speed control are shown in Figure 8. The
black line represents the control signal of the Hall vane sensor in Figure 8 (a) and (b). The red line represents thecontrol signal of the tachometer in Figure 8 (a). The tachometer used in this experiment is an optical rotation
speed sensor type, which is designed for the high-rotation-speed region (up to 300,000 min-1).
As seen in the figure, the Hall vane sensor allows controlling the air turbine spindle when the target speed is 500
min-1 even though a peak signal arises between 0 and 100 second. In particular, the peak arises when the rotation
speed is lower than 100 min-1, and the rotation control signal from the Hall vane sensor is better than the control
signal from the tachometer.
Hall vane sensor
Tachometer

Rotation speed N [min-1]

1000
900
800
700
600
500
400
300
200
100
0

0

500

1000

Time t [s](a)Rotation speed N [min-1]

1000
900
800
700
600
500
400
300
200
100
0

0

500

1000

Time t [s](b)Figure 8. Step response of the rotation speed control.
(a) Control signal of the Hall vane sensor compared with the tachometer.
(b) Hall vane sensor control signal.

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The 11th JFPS International Symposium on Fluid Power 2020

CONCLUSION
This study evaluated the air turbine spindle applied with the HPQR. This research implemented a Hall vane
sensor to compensate for the limitations of using a tachometer to measure the rotational speed and an analog
Hall effect sensor to examine its relationship with the MRF damper. The implement the Hall vane sensor
because it is a non-contact type sensor with a compact size that can be used for various purposes at a reasonablecost.
As a result, we found that the Hall vane sensor could be applied with an air turbine spindle to control therotational speed at a target speed of 500 min-1 even though some limitation arose when the speed was lower than
100 min-1. This study, therefore, showed that combining the Hall vane sensor and the tachometer could benefit
and work adequately in this type of system. Furthermore, the analog Hall effect sensor could detect the magnetic
field of the MRF damper, which is expected to be related to the viscosity measurement of the MRF in future
work.
Moreover, we will consider redesigning the position of the Hall vane sensor to develop an air turbine spindle fora milling machine to study the tool wear in the case of the low-speed condition (less than 1,000 min-1). We will
install the system following the planned diagram installation and adapting additional algorithms for further
enhancement.

ACKNOWLEDGMENTS
This research was supported by JSPS KAKENHI Grant Number 20K04203.

REFERENCES
1. KATO, T., KAWASHIMA, K., FUNAKI, T., TADANO, K., KAGAWA, T., A New High-Precision Quick Response
Pressure Regulator for Active Control of Pneumatic Vibration Isolation Tables, Precision Engineering, Vol. 34, 2010,
pp. 43-48.
2. KATO, T., A Review of HPR and Its Applications: Control of Air Turbine Spindle and Hybrid Electro-Pneumatic UltraPrecision Vertical Positioning Stage, International Conference on Fluid Power and Mechatronics (FPM), Harbin, China,
Aug. 5-7, 2015.
3. TSUKIYAMA, Y., KATO, T., MATSUO, K., Evaluation of Energy Consumption of Hybrid Electric-Pneumatic UltraPrecision Vertical Positioning Device, 10th JFPS International Symposium on Fluid Power, Fukuoka, Oct. 24-27, 2017.
4. KATO, T., HIGASHIJIMA, G., YAZAWA, T. et al., Rotation Control of Air Turbine Spindle using High Precision
Quick Response Pneumatic Pressure Regulator, Transactions of the Japan Fluid Power System Society, Vol. 45, No.1,
2014, pp. 8-14.
5. KATO, T., OTSUBO, T., SHIMAZAKI, K., et al., Tool Wear Estimation Method in Milling Process Using Air Turbine
Spindle Rotation-Control System Equipped with Disturbance Force Observer, International Journal of
Hydromechatronics, Vol. 1, No. 4, 2018, pp. 384-402.
6. OKAMOTO, Y., YAZAWA, T., KATO, T., et al., Study on Small-Diameter Ball End Mill Milling of Air Turbine
Spindle by Rotation Speed Control, 10th JFPS International Symposium on Fluid Power, Fukuoka, Oct. 24-27, 2017.
7. SRIRATANA, W., MURAYAMA, R., Application of Magnetic Field Method for Measuring Lubricant Viscosity,
Materials Science Forum, Trans Technical Publications, Switzerland, Vol. 750, 2013, pp. 116-124.
8. SATTHAMSAKUL, S., SRIRATANA, W., SRIRATANA, L., Empirical Study on Influences of Electromagnetic Field
on Hall Effect Sensor for Analysing Oil Lubricant Deterioration, 11th International Conference on Mechatronics andManufacturing (ICMM), Vol. 895, 2020.
9. LI, J., QIN, C., GUO, S., WANG, J., Magnetic circuit design and performance analysis of a rotary magnetorheologicaldamper with new structure, IEEE 8th International Conference on CIS & RAM, Ningbo, China, 2017.
10. SRIRATANA, W., MURAYAMA, R., Application of Hall Effect Sensor: A Study on the Influences of Sensor
Placement, IEEE International Symposium on Industrial Electronics, Taipei, Taiwan, May. 28-31, 2013.
11. FERNANDEZ, D., FERNANDEZ, D., MARTINEZ, M., et al, Resolver Emulation for PMSMs Using Low-Cost Hall
Effect Sensors, IEEE Energy Conversion Congress and Exposition (ECCE), Baltimore, MD, USA, Sept. 29-Oct. 3, 2019.
12. RANGAN, C.S., SARMA, G.R., MANI, V.S.V., Instrumentation Devices and Systems, 2nd edition. McGraw-Hill, New
Delhi, 2006, pp. 43-45.

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Study on Multi-Cylinder Type Wind Powered Air Compressor
Applied Hypocycloid
Ryota TANOUE* and Toshinori FUJITA***Tokyo Denki Univ. Dept. of Mechanical Engineering.
5 Senju Asahi-cho, Adachi-ku, Tokyo, 120 8551 Japan(E-mail: 20kmf13@ms.denadi.ac.jp)**
Tokyo Denki Univ. Dept. of Mechanical Engineering.
5 Senju Asahi-cho, Adachi-ku, Tokyo, 120 8551 Japan(E-mail: tfujita@mail.denadi.ac.jp)

Abstract. Air compressors are machines needed in production lines because of their versatility in compressed air, butthey have the problem of high-power consumption. The power consumption of air compressors that produce
compressed air is high, and its reduction is required from the viewpoint of greenhouse gas reduction. In this study, we
propose a wind powered air compressor that uses wind power, a renewable energy source, to drive the compressor inorder to reduce greenhouse gas emissions. In this paper, the principle and structure of the fabricated wind powered air
compressor are presented and its characteristics are measured. As a result, a maximum discharge pressure of 0.6 [MPa]and a maximum efficiency of 9 [%] were obtained at an annual average wind speed of 3 [m/s] in Japan. It was foundthat the system can be put to practical use if the efficiency is improved.
Keywords: Air compressor, Wind velocity, Flow rate, Efficiency.

INTRODUCTION
Compressed air is widely used in production lines for cleaning, conveying and so on. Today, most of thecompressed air is produced by electric air compressors. About 80 [%] of the energy consumed in industry iselectric energy. About 20 [%] of these are used in electric air compressors, which account for a large proportion.
As the reduction of greenhouse gases, which are the main cause of global warming, has become a global issue,
there is a need to reduce the power consumption of electric air compressors. For this reason, energy-saving
technologies such as inverters and number control have been introduced into electric air compressors. However,there is a limit to the amount of reduction that can be achieved [1]. For this reason, energy-saving technologies
such as inverters and number control have been introduced into electric air compressors. However, there is a limit
to the amount of reduction that can be achieved.
On the other hand, the use of renewable energy is one of the measures to reduce greenhouse gas emissions, andwind power is becoming increasingly popular due to its cost and simple structure [2]. However, wind power
generation is an unstable method of power generation in which the amount of power generated fluctuates greatly
depending on the weather. In addition, large wind turbines suffer from power transmission losses, while small
ones require inverters and other power transmission equipment, resulting in high prices.
With this background, we proposed a wind powered air compressor that uses wind turbines to power the aircompressor for more effective use of wind energy. Compared to the case where the air compressor is driven bywind power generation, this system can be introduced easily and at low cost. In addition, since it does not use
electricity, only additional piping is required when it is installed in factories. If a wind powered air compressor
can be realized, it can be installed on a vacant space such as the compressed air used, thereby reducing the numberof electric air compressors, and leading to a reduction in greenhouse gases.
This study, we proposed and developed an air compressor suitable for a wind powered air compressor, andfabricated a wind powered air compressor.

PROPOSED WIND POWERED AIR COMPRESSORUsage Example of Wind Powered Air Compressor
Fig.1 shows an example of compressed air supply in a production line in an industrial plant. Normally, an air
compressor is not installed for each machine in a factory, but several compressors are used to supply compressed air

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The 11th JFPS International Symposium on Fluid Power 2020

to the air pipes that are spread all over the factory. The wind-powered air compressor proposed in this study is notintended to supply all of the compressed air used in the factory, but to be used in conjunction with an electric air
compressor as a supplement.
The pressure fluctuates in the range of 0.6 ~ 0.9 [MPa]. Therefore, there is almost no problem even if the dischargepressure fluctuates due to changes in wind speed when a wind-powered air compressor is added. Besides comparedto the case where the air compressor is driven by wind power generation, it is obvious that the installation of a windpowered air compressor is simple and low-cost, as it only requires connection to the air piping, whereas in the case ofan electric compressor, the power supply and controller must be reinstalled when adding equipment.

Figure 1. An Example of Compressed Air Supply in a Production Line inan Industrial Plant

Figure 2. Wind Powered Air
Compressor

Structure of Wind Powered Air Compressor
Fig.2 shows wind powered air compressor. Wind powered air compressor consists of a wind turbine, a reduction gear,
and an air compressor. Therefore, the driving shaft of the wind turbine is connected to a reduction gear to amplify thetorque, and then the air compressor is driven.
Fig. 3 shows the structure of the proposed air compressor based on the hypocycloid. Unlike an electric air compressor,
a wind powered air compressor must be able to drive at low speed and low torque due to low air velocity. Screw
type air compressors are unsuitable for this driving condition. The reciprocating type has a general piston-crankmechanism that causes pivoting of the piston head motion, which worsens the sealing performance of the packing.
Therefore, we adopted a linear-crank mechanism using hypocycloid as shown in Fig. 3(a). The hypocycloid is the
trajectory drawn by a fixed point on the circumference of a moving circle that rolls while inscribed on a fixedcircle, drawing an ideal straight line when the ratio of the fixed circle to the moving circle is 2:1. Since there is no
pivoting motion, this reduces the frictional force and improves sealing performance. If this mechanism were to berealized using gears as is, the mechanism would become complicated. Therefore, as shown in Fig. 3(b), theexternal gear is replaced with a link, and a guide is attached to the piston rod to constrain the movement of theinternal gear. The linkage was extended as shown in Fig. 3(c), and cylinders were arranged with a phase shift of 90[°].
By arranging another pair of cylinders point-symmetrically to the axis of revolution, the four-cylinder system wasmade possible.

(a) Gear Type(b) Link Type2

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The 11th JFPS International Symposium on Fluid Power 2020(c) Link Type with Multi-Cylinder
Figure 3. Structure of the Proposed Air Compressor Using Hypocycloid

PROTOTYPE OF WIND POWEREDD AIR COMPRESSOR
Fig.4(a) shows the fabricated wind powered air compressor. The blades of the wind turbine were trapezoidalshaped flat blades, referring to multi-blade wind turbines, which provide high torque and high output at low
rotation speed, and 6 of these blades were used. The pitch angle of the rotor was 60[deg], the diameter was 450[mm], and the reduction ratio was 60:1. The wind powered air compressor requires a large amount of power tocompress the air. The structure of the air compressor is shown in Fig. 4(b). The bore diameter of the compressor
cylinder is 40 [mm] and the piston stroke is 12 [mm]. The packing is a low-friction pneumatic cylinder packing.

(a) The Fabricated Wind Powered Air Compressor(b) Structure of the Air Compressor

Figure 4. Fabricated Wind Powered Air Compressor

CHARACTERISTIC MEASUREMENT
Measurement of Wind Turbine Characteristics
First, the single wind turbine characteristics were measured. Fig.5 shows the measurement method of the windturbine. A pulley is attached to the output shaft of the wind turbine and a thread is applied. A weight load was
attached to one end of the thread via a pulley, and a scale to the other end. With a constant and uniform windspeed generated in the wind tunnel, the weight load was added in 0.15 [kg] increments until the wind turbine
stopped. The number of rotations at this time was measured using an encoder, and the torque generated wasmeasured by multiplying the indicated value on the scale by the pulley radius. The wind turbine efficiency was
also calculated from the wind energy, measured torque and revolution speed. The measurement results for windspeeds of 3[m/s], 4[m/s] and 5[m/s] are shown in Fig.6. The theoretical maximum efficiency of wind turbines is59.3%, and those with 40% efficiency are used for wind power generation. Compared to these, the efficiency
obtained was in the low 20% range, which was not good.

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Torque

[Nm]

0.15

Scale
Threadwind speed[m/s]experimentapproximate

3

4

5

0.1
0.05

Pulley0[%]Wind turbine

Efficiency

Load
TachometerWind tunnel

30
20
10
0

Figure 5. The Measurement Device for Wind Turbine
Characteristics

200
400
Revolution [rpm]600

Figure 6. The Measurement Results of Wind Turbine
Characteristics

Measurement of Air Compressor Characteristics
The characteristics of the air compressor were investigated in the same way, and the measurement apparatus isshown in Fig.7. The air compressor was revolved at a constant speed by a motor, and the required torque of thecompressor was measured by a torque sensor. An isothermal pressure vessel with a volume of 300[cc] wasinstalled at the discharge side of the compressor, and the discharge pressure was read by a pressure sensor. The
discharge flow rate was obtained from the equation of state of the gas by differentiating the pressure response of
the isothermal pressure vessel. The product of the torque sensor value and the revolution speed was used as the
input power, and the exergy of the compressed air was obtained from the discharge pressure and flow rate, andthe efficiency was calculated as the output power.
Fig.8 shows the measured pressure flow characteristics and efficiency when the compressor is revolved at a low
speed of 5[rpm], as expected in an actual wind turbine compressor. The maximum discharge pressure of the aircompressor was 0.6[MPa], and the maximum efficiency was 60[%].
In general, small reciprocating air compressors are said to have about 30[%]. This value includes the motor
efficiency, but the motor efficiency is more than 80[%], so even if the motor efficiency is not included, we cansay that we have achieved a compressor with a very high efficiency compared to conventional air compressors.

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Discharge Pressure[MPaG]

The 11th JFPS International Symposium on Fluid Power 2020

Motor

Air Compressor

0.8
0.6
0.4
0.2
0

Pressure sensor

60

Isothermal Chamber

Data Logger

Efficiency

Torque Sensor

[%]

P

M

40

20

0
0

Figure 7. Measured Characteristics of Air Compressor

0.1
Flow Rate [NL/min]

0.2

Figure 8. Measuring Apparatus for Air Compressor
Characteristics

Measurement of Wind Powered Air Compressor Characteristics
Fig.9 shows the apparatus used to measure the characteristics of the wind powered air compressor. A constantwind speed was given to the wind powered air compressor by a blower and a wind tunnel, and the compressed air
generated at that time was filled into an isothermal pressure vessel. From this pressure response, the pressure flow
characteristics were obtained as in the measurement of the air compressor itself. As for the wind speed, the averagewind speed in Japan is said to be 3[m/s] [3], so the wind speed was set to 3[m/s] and 5[m/s] for the measurement.
Fig.10 shows the characteristics of the obtained wind powered air compressor. The maximum discharge pressure
was 0,6 [MPa]. The maximum discharge pressure was 0,6 [MPa], which is slightly lower than the maximum
discharge pressure of general air compressors, which is 0,6 to 0,8 [MPa], but still practical. On the other hand, it
can be seen that the flow rate increases as the wind speed increases, but the efficiency decreases. This means that
it is necessary to match the wind turbine characteristics with the air compressor characteristics so that the
maximum efficiency points are the same.
Consider the case where an air compressor is driven by electricity generated by a wind turbine in a remote location.
If we assume that the efficiency of the wind turbine is 40[%], the transmission loss of electricity is 20[%], and theefficiency of the compressor is 50[%], the overall efficiency is 16[%]. The maximum efficiency obtained is lower
than this value, and it can be said that the efficiency needs to be improved. Since the efficiency of a single windturbine is low, we believe that it will be possible to improve the efficiency of a single wind turbine and to matchthe characteristics of wind turbines and air compressors.

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0.8

wind velocity[m/s]3experiment

5

Wind

Wind Powered
Air Compressor

Delivery Pressure

0.6

Pressure Sensor

0.4
0.2
0

P
10
[%]

Data Logger

Efficiency

Isothermal Chamber

50Figure 9. Apparatus Used to Measure the Characteristics of the Wind
Powered Air Compressor

0.1
0.2
0.3
Flow Rate [NL/min]Figure 10. overall efficiency (𝑢=3,5[m/s])

CONCLUSION
We proposed and fabricated a prototype of a wind powered air compressor that utilizes wind energy to reducegreenhouse gas emissions. A compression mechanism using a hypocycloid was shown for that air compressor.
The characteristics of the air compressor itself were measured and showed a high efficiency of 60[%]. As a resultof measuring the characteristics of the prototype wind powered air compressor, a maximum discharge pressure of
0.6[MPa] and a maximum efficiency of 9[%] were obtained at the annual average wind speed of 3[m/s] in Japan.
It was found that the discharge pressure was about the same as that of an ordinary air compressor, and if theefficiency could be improved, it could be put to practical use.

REFERENCES
1. ONEYAMA, N., Energy conservation in pneumatic systems, Energy Conservation Center, 2003, pp.9-12.
2. USHIYAMA, I., Introduction to Wind Turbine Engineering -From Basic Theory to Wind Power Generation
Technology-(2nd edition), Morikita Publishing, 2004,pp1-5,139-140,156-162.
3. https://www.data.jma.go.jp/obd/stats/etrn/view/monthly_s3.php?prec_no=14&block_no=47412&year=2019&month=&day=&view=a (accessed on 1 May, 2021)

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Development of Bidirectional Arm Curl Machine Using Pneumatic
Artificial Rubber Muscles
Toshihiro KAWASE*,**, Tomoya NAKANISHI*, Shintaro YOSHIDA***, Shingo OHNO***,Ryo SAKURAI***, Tetsuro MIYAZAKI****, Takahiro KANNO*****, Maina SOGABE****,
Yoshikazu NAKAJIMA*, Kenji KAWASHIMA*****

Institute of Biomaterials and Bioengineering, Tokyo Medical and Dental University
2-3-10 Kandasurugadai, Chiyoda-ku, Tokyo, 101-0062 Japan(E-mail: kawase.bmi@tmd.ac.jp)**
Institute of Innovative Research, Tokyo Institute of Technology
4259 Nagatsuta, Midori-ku, Yokohama, Kanagawa 226-8503 Japan***
Bridgestone Corporation
3-1-1 Ogawahigashi-cho, Kodaira-shi, Tokyo 187-8531 Japan****
Department of Information Physics and Computing, The University of Tokyo
7-3-1 Hongo, Bunkyo-ku, Tokyo, 113-8656 Japan*****
RIVERFIELD, Inc.
Yotsuya Medical Bldg. 5th floor, 20 Samon-cho, Shinjuku-ku, Tokyo, 160-0017 Japan
Abstract. Physical training with unpredictable perturbations is commonly used to prevent injuries of athletes orolder adults. In our previous study, we have developed an arm curl machine that exerts unpredictable loads using apneumatic artificial rubber muscle (PARM). In this study, we extended the arm curl machine to exert load with
unpredictable direction on an elbow. The machine was extended to generate both of flexional and extensional forces
with antagonistic actuation of PARMs. Force control of the PARMs was implemented for accurate control of thetension of PARMs. The effect of training using the extended arm curl machine was tested by an experimental
training with six healthy participants. The results suggested that the training using load with unpredictable direction
generated by the machine had effects on the reduction of extra muscle activation on trunk.
Keywords: Muscle training, Pneumatic artificial rubber muscle, Antagonistic actuation

INTRODUCTION
Physical training with unpredictable perturbations is commonly used to prevent injuries of athletes or older
adults [1][2]. In our previous study, we have developed an arm curl machine that exerts unpredictable loads
using a pneumatic artificial rubber muscle (PARM) and showed the possibility of effective muscle strength
improvement against unpredictable load [3]. In the arm curl machine, the direction of exerted force was only
downward as in conventional arm curl machines. In real situations in exercises, the direction of perturbations
may also be unpredictable. Generally, the direction of load during movement or posture maintenance is a key
factor that determine the pattern of coordinated activity of muscles. Incorrectly coordinated muscle activities
under the load may cause dangerous postural change, and training to prevent the hazard, which is caused by loadwith unpredictable direction, should be considered. In this study, we extended the arm curl machine to exert
flexional and extensional force on an elbow with antagonistic actuation using PARMs and confirmed its effectin a training using unpredictable loads by an experiment.

BIDIRECTIONAL ARM CURL MACHINE
Figure 1 shows the developed arm curl machine. The weights of a commercially available arm curl machine(Freemotion Epic Bicep, Icon Health & Fitness, Inc.) were disconnected from its joint, and PARMs (supplied by
Bridgestone Corp.) were connected to the joint by wires on flexion (upward) and extension (downward) sides.
We used double PARMs on the flexion side for gravity compensation. A rotary encoder (MES-12-1000PST8E，
MICROTECH LABORATORY) was attached on the joint of arm curl machine to calculate gravity force on themachine. A force sensor (FGP-50 and FGP-100, Nidec-Shimpo Corp.) was attached to each set of the PARMs
and a servo valves (MPYE-5-1/4-010-B, Festo) was connected to each PARM for the control of generated
forces.

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(a)

(b)

Figure 1. The developed bidirectional arm curl machine. (a) Appearance of the machine. (b) System overview.

The block diagram of the control system is shown in Fig. 2. Force control helped by the force sensors was used
to control the tension of PARMs accurately, while the previous version used pressure control [3]. The gain
parameters for the force control are shown in Table 1. A PARM on the extension side exerted extensional loads,
and two PARMs on the flexion side exerted flexional loads and torque for gravity compensation (Fig. 3). This
control system could supply constant force regardless of the posture.

Figure 2. Block diagram of the arm curl machine
Table 1. Gain parameters for force control of PARMs

Kep [V / N]
Ked [V / (N s)]
Kfp [V / N]
Kfd [V / (N s)]

0.11
0.00010.09
0.0001Figure 3. Force on the machine and a trainee

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EXPERIMENTThe arm curl machine could present loads with unpredictable direction (extension or flexion), which could not bepresented by the previous version. To confirm its effect on physical training, we conducted an experiment withparticipants.
Six healthy males participated in the experiment. First, the participants maintained their posture under a 1minute load pattern with randomized amplitude, direction and timing (pre-test task). Second, they trained their
posture maintenance under a load pattern with unpredictable direction (Direction pattern; Fig. 4(a)) (trainingtask). The training duration was 3 minutes in total for each participant. Finally, the participants carried out thefirst 1-minute task again (post-test task). The series of the tasks was conducted with other three load patterns
during training without unpredictable direction: load patterns with unpredictable amplitude (Amplitude pattern;
Fig. 4(b)), with unpredictable timing (Timing pattern; Fig. 4(c)) and with predictable, cyclic change (Cyclic
pattern; Fig. 4(d)).
Through the experiment, we measured surface electromyography (EMG) signals of four muscles: arm muscles(biceps and triceps) and trunk muscles (rectus abdominis and latissimus dorsi). We used EMG measurement
system with active electrodes (Bagnoli EMG System, DELSYS) and calculated EMG amplitudes duringflexional and extensional load in pre-test and post-test tasks by root-mean-square value of EMG signals. We
compared the ratio of EMG amplitude in the post-test task to the amplitude in the pre-test task as an indicator of
the change of muscle activity.

(a)

(b)

(c)

(d)

Figure 4. Example of reference force of the PARMs during a training task with four load patterns: Direction pattern (a),Amplitude pattern (b), Timing pattern (c) and Cyclic pattern (d).

RESULT
Figure 5 shows the ratio of EMG amplitude in the post-test task to the amplitude in the pre-test task under the
training with four load patterns (means and standard deviations between the participants). In the case of training
with Direction pattern (Fig. 5(a)), the EMG amplitude of rectus abdominis showed significant reduction after the
training (t-test, p < 0.05) in both direction of the load. This phenomenon could not be found in the other training
tasks (Fig. 5(b-c)). The rectus abdominis is the muscle that contributes to posture maintenance of upper body;
however, in the situation of use of arm curl machine, the activation of rectus abdominis would not be required
because the upper arm is fixed on the machine. The results suggest that the training task with unpredictabledirection could contribute to restrain over-activation of trunk muscles under unpredictable perturbations.

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The 11th JFPS International Symposium on Fluid Power 2020

(a)

(b)

(c)

(d)

Figure 5. Ratio of EMG amplitude in the post-test task to the amplitude in the pre-test task under the training with four load
patterns: Direction pattern (a), Amplitude pattern (b), Timing pattern (c) and Cyclic pattern (d). Means and standard
deviations between the six participants are shown. The asterisks show the significant difference of the ratio from 100% (p <0.05).

CONCLUSION
We developed an arm curl machine using PARMs that could exert extensional and flexional loads. We showedthat the machine could be used for a training with loads with unpredictable direction, and the training could
suppress extra muscle activation under unpredictable perturbations.

ACKNOWLEDGMENTS
This work was supported by the Bridgestone Corporation and KAKENHI (20K19594, 21H04544). This
research was also based on the Cooperative Research Project of the Research Center for Biomedical
Engineering.

REFERENCES
1. PATERNO. M. V., et al., Neuromuscular Training Improves Single-Limb Stability in Young Female
Athletes, Journal of Orthopaedic & Sports Physical Therapy, 2004, 34-6, pp.305-316.
2. MANSFIELD. A., et al., Effect of a Perturbation-Based Balance Training Program on Compensatory
Stepping and Grasping Reactions in Older Adults: A Randomized Controlled Trial, Physical Therapy, 2010,
90-4, pp.476-491.
3. NAKANISHI, T., et al., Development of an Arm Curl Machine with Variable Resistance Using Pneumatic
Artificial Rubber Muscle, Proceedings of the 2019 IEEE/RSJ International Conference on Intelligent Robots
and Systems, 2018, pp.3830-3835.

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Design of a Pneumatic Oscillator for Paper Machine's Doctor Blade
Systems
Michele Gabrio ANTONELLI*, Pierluigi BEOMONTE ZOBEL*, Massimiliano
CENTOFANTI**, Stefano COLAIUDA***

Department of Industrial and Information Engineering and Economics, University of L’Aquila
P.le Pontieri, Monteluco di Roio, L’Aquila, 67100 Italy(E-mail: pierluigi.zobel@univaq.it)**
SMC Italia SpA
Località Recocce, Via delle Industrie, Carsoli (AQ), 67061, Italy

Abstract
In continuous paper machine, the cleaning of the surfaces of the cylinders, adopted for the transformation process ofthe paper, is essential. The component intended for this scope is the Doctor Blade. It is a steel blade that extends forthe total length of the cylinder surface and that can move towards the cylinder and along it. This second motion,transversal with respect to the machine direction, is oscillatory. Hydraulic, pneumatic, and electro-mechanicalactuators, commercially available, are typically adopted to carry out this motion. In the present work, a pneumaticoscillator, to be mounted on one end of the Doctor Blade is described. It is made of a pneumatic cylinder equippedwith two integrated limit switches. The proposed device was conceived to guarantee the following performances: low
cyclic motion, simple and intuitive speed adjustment, ease of assembly and feeding, shielding from external corrosive
agents.
Keywords: Pneumatic Oscillator, Actuator Design, Paper Machine, Doctor Blade

1. INTRODUCTION
Pneumatic systems are particularly suitable in environments with moisture problems or the presence of flammablematerial, in which electrical components are not reliable, and for their low cost and their good power to weightratio. One of the sectors which requires the use of pneumatic actuators is the paper production sector.
The paper production begins with the supplying of wood from trees, continues with the production of paper coils
through the continuous machine with the processing of different cellulose mixtures and ends with thetransformation of these large mother coils into finished products.
Neglecting the transformation cycle from the starting wood until the produced cellulose mixtures, once transferredto the paper mill, the latter are spread with water and undergo purging processes, so as to eliminate possible
polluting residues, and refining, to give certain mechanical characteristics to the product [1]. In the entrance tothe continuous machine there is a suspension of water and fibrous material (concentration less than 1%) fromwhich the paper will be produced. Along the continuous machine, the suspension gradually undergoes atransformation process that consists in a progressive drying, which takes place by a sequence of operations:forming by natural drainage and suction, pressing, evaporation [2] to obtain paper coils with humidity around 5%[3]. In order to perform the sequence of operations, the continuous machine is made of several components: theheadbox for conveying the fibrous dough, diluted over 99% of water, into the forming area where the dough, once
released on training canvas moved by cylinders, undergoes, in the form of sheet, the phenomenon of drainage inorder to remove part of the water from it; presses, made of some cylinders covered by felt, for reducing the residual
amount of water; a dry zone, made of the Yankee cylinder and a metal hood, for evaporating the water from 5060% until a content of about 5%; finally, a cylinder for winding the paper coils until they reach the desired
diameter. Coils represent the raw material of the paper works, through which the finished products are produced
and packed.
Among the finished products, tissue papers are intended for hygienic and sanitary use with the principal
characteristic of softness and bulkiness. These properties are given to the paper thanks to some factors as thedegree of adhesion of the paper sheet to the forming cylinders and the difference in speed between the dryer
cylinder and the coiler. Another important factor is the action of the Doctor Blades (DBs), steel blades positioned
in contact with the cylinders, which have the main tasks of cleaning the surfaces of the cylinders and detaching
the sheet leaving a series of small waves imprinted on it [4]. One aspect to be considered is the proper position of
a DB with respect to the cylinders that can be adjusted according two modes [5]. The first mode expects the
adoption of pneumatic or hydraulic cylinders that approach and move away the DB and its support. The second

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mode expects the adoption of a couple of alternatively deformable pneumatic tubes positioned on the sides of ahinge: as pneumatic muscles [6, 7], the radial expansion of the tubes provides for a pushing action on the supportof the DB that rotates around the hinge, approaching or moving away the cylinders. Another aspect is theoscillatory motion of the DB (forward and backward) along a cylinder for scraping the external surface of thecylinder. It can be carried out by hydraulic, pneumatic, and electro-mechanical actuators [5]. Pneumatic ones aremost common: they can be made by one or more pneumatic air springs with slide mechanism [5, 8] or bypneumatic cylinders [9, 10]. Registered innovations of DB are: new materials as hard carbon and a speciallydeveloped epoxy and the equipment of sensors for detecting incorrect settings of the DB [11]; simpler techniqueof disassembling for a regular maintenance of the pressure tubes, a regular cleaning of the blade pocket and a
regular inspection of the top-plate [12]; material water repellent to facilitate cleaning operations while ensuring
longer life due to absence of protective treatments [13].
The present research is focused on the development of a pneumatic oscillator to be mounted on a commercial
continuous paper machine, for the oscillatory motion of DBs along cylinders. The proposed pneumatic oscillator
is made of a pneumatic cylinder equipped with two limit switching systems installed within the two heads of thecylinder. Each system is made of a 2/2 monostable pneumatic valve directly commanded by the piston of thecylinder. Such pneumatic oscillator is mounted at one end of the support of each DB in a continuous machine forthe paper production. The concept idea and the design of the pneumatic oscillator are here reported.

2. BACKGROUND
2.1 The doctoring system on the continuous paper machine
An effective and hassle-free cleaning system is essential for the operation of the modern, automatic, and computercontrolled paper machine. The use of good cleaning practices and the best available materials can result in a bettermachine drivability, safer operations, reduced water consumption, energy savings and an excellent quality of thepaper. A very important part of the continuous paper machine is the system of scrape of the cylinders of theforming and pressing processes, made of oscillating blades called DBs. As shown in Figure 1, a DB system expects
a blade, that extends for the total length of the cylinder surface, a support beam of the blade equipped with a tankfor the containment of processing fluids and a positioner of the support beam. When it is required, the entire
system must oscillate along the transverse direction of the machine.

Figure 1. Doctor Blade system with three blades

The DB system has:
- to ensure the proper cleaning of the surface of the cylinder, avoiding the accumulation of foreign material on it,
which can compromise the quality of the paper sheet and damage the cylinder;- to avoid a sanding through the surface of the cylinder and maintain the proper surface roughness for a very long
period;- to ensure the proper detachment of the paper sheet during the forming and pressing process;- to redirect the paper sheet and prevent it from wrapping around the cylinders.

2.2 Details of the application
The DB system of interest is shown in Figure 2: it is possible to distinguish the blade, the tank for the fluids andthe flange interface between the pneumatic oscillator, to be mounted, and the support beam.

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Figure 2. Details of the commercial Doctor Blade system to be moved

The technical requirements are reported in the follow:- the amplitude of the oscillatory motion can vary in the range 10 - 20 mm;- the frequency of the oscillatory motion can vary in the range 1 - 10 cycles/min;- the oscillatory motion has to be linear and fluid.
It means that the cylinder of the oscillator must move at low adjustable speeds. Hence, it must be equipped withan integrated adjustment device. Finally, a low friction pneumatic cylinder (equipped with proper gaskets andgreases) must be adopted in order to minimize the stick-slip phenomenon. In addition, no significant backslash in
the drive mechanism must be occurred. About the continuous functioning, the pneumatic oscillator must be an
automated device. For this reason, it must be equipped with only one inlet port for the compressed air: once
connected to the pneumatic line, the oscillator must operate autonomously. Finally, since the oscillator works ina corrosive environment and with a high humidity rate, the external components must be made of stainless steel
AISI 304.
The technical specifications of the cylinder of the pneumatic oscillator are reported in Table 1.

Force

Table 1. Pneumatic cylinder technical specifications
11 kN at 6 bar

Stroke20 ± 1 mm

Number of cycles

1 – 10 cycles/min

Operating pressure

2 ÷ 8 barIntegrate regulation system

Single air inlet port and a flow regulator for speed adjustment

Suitable for corrosive environments

Material of the outer components: AISI 304

3. MATERIALS AND METHODS
3.1 Preliminary consideration to the design of the pneumatic oscillator
The first step was the definition of the regulation circuit necessary to allow the oscillation of the cylinder. The
type of the circuit influences the architecture of the cylinder. Two different solutions were preliminary conceived:
the first one is based on timed control pulses of the pilot valve; the second is based on mechanical limit switch
valves.
The first solution requires the development of the circuit shown in Figure 3, easily to be integrated on iso/standard
cylinder and with a low cost of production. With reference to the Figure, flow regulators S1 and S2 are necessary
for the speed adjustment; the flow regulators S3 is connected at one end to the line of the chamber (a) and at theother to the command side of the pilot valve for powering the chamber (b). The current configuration of the circuit
expects that compressed air moves to the chamber (a) and provides for the leftward motion of the piston with aspeed depending on the throttling level of the regulator S1; moreover, compressed air moves through the flow
regulator S3 and after a certain time, depending on the throttling level of the regulator, gives the command input(command 14) for switching the pilot valve. When it occurs, both the lines to S 3 are connected to the exhaust andthe piston moves rightward. Then the flow regulator S 4 works as S3, but on the opposite side of the pilot valve,giving the timed pulse command input 12.

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Figure 3. Regulation circuit for the pneumatic oscillator: solution 1

Dimensions of the flow regulator guided the choice to replace the pneumatic timer, bulkier. This solution is timebased and allows the cylinder to move autonomously. Preliminary experimental tests confirmed the describedworking principle. Nevertheless, to assure the same behavior of the motion along the rightward and leftward
stroke, the length of the pneumatic lines should be equal, and the throttling level of the flow regulators should be
the same. Moreover, with the change of the desired speed, four manual regulations must be carried out. In addition,synchronization between the speed of the piston and the switching time must be carried out. Finally, due to thelow value of the switch pressure of the pilot valve (about 0.03 MPa), flow regulators with a narrow passage section
are necessary.
For these critical aspects, the second solution was explored. It requires the development of a circuit that adopts
two limit switch valves. With reference to Figure 4, the switch valves a 0 and a1 are mechanical operated 2/2
monostable pneumatic valves: the pilot valve, a 5/2 monostable pneumatic valve, normally provide for theleftward motion of the piston; when the left rod of the cylinder activates a0, the latter provides for the 14 command
input for switching the pilot valve; hence, the piston rightward moves and the pneumatic line between a 0 and thepilot valve remains pressurized; when the right rod of the cylinder activates a 1, the pilot valve switch to its normal
position and the pneumatic line between a0 and the pilot valve is connected to the exhaust. In order to have aunique point of adjustment of the speed, only one flow regulator is adopted: it is connected to the exhaust portsof the pilot valve.

Figure 4. Regulation circuit for the pneumatic oscillator: solution 2

In preliminary experimental tests, 2/2 monostable valves were replaced by 3/2 monostable valves, in which theexhaust ports were capped. For a power supply pressure equal to 0.65 MPa, circuits confirmed the proper
functioning. The adjustment of the throttling of the flow regulator was easy to carry out with an improved behavior
with respect to the first solution.

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3.2 The proposed solution
Figure 5 shows a section view of the cylinder. It is possible to distinguish the two limit switches and how thepiston touches and activates the internally integrated valves: if the CHAMBER (A) is powered and the
CHMABER (B) unloads, the piston moves downward and touches the limit switch of the lower head. The force
exerted on the rod of the limit switch causes air to pass through and the pilot valve, mounted externally of thehigher head, to be switched. When the pilot valve switches, the CHAMBER (B) is powered, the CHAMBER (A)unloads and the piston moves upward. When the other limit is operated, the cycle is repeated.

Figure 5. Section view of the cylinder and detail of the two integrated limit switches

The limit switch system is the most important element of the oscillator and is shown in Figure 6. It consists of a
movable rod, kept in a resting position by two steel springs EN 10270-1 SH(C) placed in parallel, which moves
inside a seat obtained directly in the two heads. The rod is characterized by two lip seals, which isolate the valve
body from the inner chamber of the cylinder, and a piston seal that isolates chamber (1) from chamber (2). The
system was designed to replicate the operation of a mechanical operated 2/2 (NC) monostable valve: when thepiston touches the rod, pushes it and the piston seal downward allowing the passage of air for the switching of thepilot valve.

Figure 6. Details of the structure of the limit switch system integrated in the cylinder heads

With reference to Figure 6, the critical area turns out to be the passage section of the fluid, from chamber (1) tochamber (2). This section must be open and closed whenever the piston touches the rod to determine the switchingof the pilot valve. The closing cap solved this problem. It was designed not only with the ultimate aim ofhermetically closing the system from the outside, but also with the aim of making possible the variation of thesection diameter in the passage between the chamber (2) and the chamber (1) and of isolating, between them, the5

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two chambers. For the latter aim, there is a dynamic seal on the rod: it is not a simple O-ring, but a piston seal.
Once mounted, the cap constitutes the chamber (1).
The chamber (1) of the limit switch mounted in correspondence of the CHAMBER (B) is always pressurized: it
is directly connected to the main pneumatic line which also provides for the power supply of the pilot valve.
During the motion of the rod of the limit switch, due to the exerted force of the piston, the passage section between
the chamber (1) and the chamber (2) increases. With reference to Figure 5, as regards the limit switch mounted incorrespondence of the CHAMBER (B), the air moves towards the command port of the pilot valve, in the formof a positive command: when the value of the pressure reaches the pilot pressure value, the pilot valve switches,
and the piston moves upward. As regards the limit switch mounted in correspondence of the CHAMBER (A), theair moves toward the exhaust port in the form of a negative command, removing the air along the command lineof the pilot valve that reaches its normal position. A silencer provides to reduce the noise due the exiting air.
The adjustment system is integrated into the cylinder through a combination of internal and external ducts. In fact,
it is necessary to replicate the pneumatic circuit shown in Figure 4 and to integrate it in the cylinder, making thepneumatic cylinder and the adjustment circuit a single and compact body. The main pneumatic branches are shownin the rendered Figure 7a, with different colors:- in light blue, between the two red lines, the supply branch allows air to be brought from the external supply line
to the pilot valve;- in light blue, in the lower part, the supply/unloading branches of the CHAMBERS (A) and (B);- in green, the control branch that brings the switching pulses from the integrated limit switching system in thetwo heads up to the pilot valve, command side;- in red, the exhaust branch of the two CHAMBERS (A) and (B) and the connection to the pilot valve exhaust
ports with the flow regulator (not visible).

(b)

(a)
Figure 7. Definition of the connecting branches of the regulation system

The designed pneumatic oscillator has a construction architecture typical of normal double-effect cylinders withthrough rod. In the Figure 7b, the cylinder looks like a compact body, in which only one head is visible externally(the one in contact with the external environment). To avoid all the components made of stainless, it was thought
to make a watertight protective casing. The head that interfaces both with the machine, by coupling flange, andwith the main line of compressed air for the cylinder supply was also conceived to be made of AISI 304, as wellas the rod that couples with the DB system to be moved.

4. CONCLUSIONSIn the present paper, the design of a new pneumatic oscillator for a Doctor Blade system was presented. The
oscillator is made of a double-effect cylinder with through rod equipped with an internally integrated limit switch
in each head. The proposed design will provide for an autonomous pneumatic oscillator with a single air inlet port
and a flow regulator for speed adjustment. The next step of the development activity expects the prototyping of
the proposed solution and a first campaign of experimental tests to validate the working principle and to achieve

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the performances of the overall system. Improvements could be introduced in order to reach and industrial andmarketable product.

REFERENCES
1. DE ASSIS, T., REISINGER, L. W., PAL, L., PAWLAK, J., JAMEEL, H., GONZALEZ R. W., Understanding the Effectof Machine Technology and Cellulosic Fibers on Tissue Properties – A Review, BioResources, Vol.13, No.2, 2018,
pp.4593-4629.
2. AXELSSON, B., Pulp for high absorption tissue products, Paper Technology, Vol.42, No.7, 2001, pp.24-26.
3. KULLANDER, J., NILSSON, L., BARBIER, C. Evaluation of furnishes for tissue manufacturing; suction box dewateringand paper testing, Nordic Pulp & Paper Research Journal, Vol.27, No.1, 2012, pp.143-150.
4. BERNI, A., MENNIG, M., SCHMIDT, H. Doctor Blade. In: Aegerter M.A., Mennig M. (eds) Sol-Gel Technologies for
Glass Producers and Users. Springer, Boston, MA, 2004, pp. 89-92.
5. https://www.valmet.com/globalassets/media/downloads/white-papers/process-improvements-and-arts/wpp_doctoring.pdf(accessed on 1 June, 2021)
6. ANTONELLI, M. G, BEOMONTE ZOBEL, P., DURANTE, F., RAPARELLI, T. Numerical modelling and experimental
validation of a McKibben pneumatic muscle actuator, Journal of Intelligent Material Systems and Structures, Vol.28,
No.19, 2017, pp.2737-2748.
7. DURANTE, F., ANTONELLI, M. G., BEOMONTE ZOBEL, P., RAPARELLI, T. Development of a Straight Fibers
Pneumatic Muscle, International Journal of Automation Technology, Vol.12, No.3, 2018, pp. 413-423.
8. WAKE, H. D., Oscillating Doctor-Blade Mechanism, United States Patent 3,750,228, August, 7, 1973.
9. https://www.valmet.com/media/articles/all-articles/valmet-doctor-oscillators-for-reliable-maintenance-free-service/(accessed on 1 June, 2021)
10. https://dcf.kadant.com/it/prodotti/raschiatura/oscillatore-varimatic (accessed on 1 June, 2021)
11. https://voith.com/corp-en/papermaking/doctor-and-creping-blades.html (accessed on 1 June, 2021)
12. https://www.clouth.com/en/ticker/2020/06/23/clouthr-topslide-the-innovative-doctor-blade-holder-with-patented-quicktop-system (accessed on 1 June, 2021)
13. https://www.flexofem.com/en/doctor-blade-system/ (accessed on 1 June, 2021)

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Genaral Session | Oil hydraulics

Components and Systems 2[GS5-2-01] Leakage Characteristics of a 3-port Pressure Reducing Valve〇Haroon Ahmad Khan1, Byeong-Il Choi2, Jung-Ho Park2, So-Nam Yun2 （1.University of
Science and Technology, Korea, 2.Korea Institute of Machinery and Materials）[GS5-2-02] Generation Mechanism of Flow Force Acting on Spool Valve〇Fumio Shimizu1, Kazuhiro Tanaka1 （1.Kyushu Institute of Technology）[GS5-2-03] The Effect of the Spline Coupling on the Rotating Assembly Tilt Behavior
in a High-speed Axial Piston Pump
〇Haogong Xu1, Junhui Zhang1, Weidi Huang1, Bing Xu1, Fei Lyv1, Xiaochen Huang1 （1.StateKey Laboratory of Fluid Power and Mechatronic Systems, Zhejiang University）[GS5-2-04] First-order Trajectory Sensitivity Analysis of Multi-level Pressure
Switching Control System
Jing Yao1,2,3, 〇Pei Wang1, Xinhao Li1, Yuwang Cheng1, Xiaoming Cao1 （1.School of
Mechanical Engineering, Yanshan University, 2.Key Laboratory of Advanced Forging&Stamping Technology and Science of Ministry of Education of China, 3.Hebei Key Laboratoryof Heavy Machinery Fluid Power Transmission and Control）[GS5-2-05] A Study on the Pulse Analysis and Vibration Characteristics of Hydraulic
System for Prediction of Check Valve Behavior〇Jeong-Woo Park1,2, So-Nam Yun1, Young-Bog Ham1, Eun-A Jeong1 （1.Korea Institute ofMachinery &Materials, 2.Chungnam National University）[GS5-2-06] Research on an Oil-hydraulic Component to Reduce Pressure Pulsation〇Yasuo Sakurai1, Misaki Hashimoto2, Moritaka Maehara3, Norikazu Hyodo4 （1.Ashikaga
University, 2.MITSUBA Corporation, 3.SAWAFUJI ELECTRIC CO., LTD., 4.TOKYO KEIKI INC.）©The Japan Fluid Power System SocietyGS5-2-01

The 11th JFPS International Symposium on Fluid Power 2020

Leakage Characteristics of a 3-port Pressure Reducing Valve
Haroon Ahmad KHAN*, Byeong-Il CHOI**, Jung-Ho PARK** and So-Nam YUN***

University of Science and Technology
217 Gajeong-ro, Gajeong-dong, Yuseong-gu, Daejeon, Korea(E-mail: ysn688@kimm.re.kr)**
Korea Institute of Machinery and Materials,
156 Gajeongbuk-ro, Jang-dong, Yuseong-gu, Daejeon, Korea

Abstract. Leakage in hydraulic system is directly related to power losses in the system which is undesirable in any
application. Pressure reducing valves have two ports or three ports depending on application. Leakage study on oilhydraulic pressure reducing valves cannot be found in prior research work which makes this study quite significant.
An experimental equipment was setup for testing the valve that included a data acquisition system and pressure
sensors. The test results were recorded and analyzed in this study. The leakage of the valve was measured in a steadystate with a stop watch and a beaker with volume markings. Instead of a flow sensor, a more basic approach ofvolume per time was followed because of the inaccuracy of flow sensors at extremely small flow rates. The valve
was actuated by solenoid with an ability of having different output pressure. The leakage characteristics wereobtained for different output pressures.
Keywords: Pressure Reducing Valves, Hydraulic Leakage, Pressure Control Valves, Solenoid Valves

INTRODUCTION
For any hydraulic system leakage is undesirable as it is a cause of power loss. Less leakage is always desirable
but leak-free systems are mostly impossible to design in certain hydraulic valves. In some cases, leakage isinevitable if two surfaces are to slide against one another as is the case with spool valves. In a pressure reducing
valve, the outlet pressure is controlled by the movement of spool inside the valve sleeve. If pressure is higher
than desired pressure, the spool pushed on one side, if it is less than desired then the spool is pushed to the otherside. So, using a spool is inevitable in such a valve and it requires some clearance between the spool and sleeve
for the sliding movement. Due to such clearance, there a small amount of flow rate that is leaked all the time.
Such a flow can be characterized as Poiseuille flow in an annular section whose formula was derived by [1] asshown in equation (1).

G
( R  R1 ) 2
4
4[ R2  R1  2]
8
ln( R2 / R1 )2

Qleak 2

(1)

Where G is the pressure drop per unit length, R1 is the inner radius and R2 is the outside radius. The power loss
due to leakage can be determined as:

Ploss,leak  PoutQleak(2)

Where Pout is the pressure obtained at the reducing valve outlet. Although this power loss may be less ascompared to losses in other parts of the hydraulic system, it still contributes to higher energy loss if operated fora long time.
Leakage can occur during steady state of the valve system or during dynamic state. In this study only steadystate leakage was studied. For this purpose, the valve settings were kept constant during the leakage
measurements as will be explained in the next sections.

VALVE DETAILSThe valve used in this study was a pressure reducing valve. Pressure reducing valves can have 3-ports or 2-ports
as shown in figure 1:

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Figure 1. 2-port and 3-port Pressure Reducing Valve [2]The advantage of using a 3-port pressure reducing valve is that it can allow the pressure at the outlet to drop by
connecting the outlet to the reservoir port of the valve. However, a 2-port valve is only used in situations where
the pressure at the outlet can never go over the limit set by the spring or the solenoid. In this study a 3-portpressure reducing valve was used and the leakage from outlet port to the reservoir port was tested.

EXPERIMENTAL TESTING
The valve was installed in a hydraulic power unit capable of providing a supply pressure of 50 bars. The inlet
port of the valve was connected to the 50 bar pressure. The outlet of the valve was blocked so that leakage flow
can be easily measure. A beaker with volumetric markings was used to collect the fluid coming out through thereservoir port. 100ml of liquid was collected and time duration was noted using a stopwatch.

Leakage Flow
Collector
Power Unit

Valve Block
Figure 2. Experimental Details of the Leakage Measurement Equipment used for this study

The details of the hydraulic equipment used in this study is shown in figure 2. The accumulator was used at theoutlet of the pressure supply from the pump in order to reduce the fluctuations in the inlet pressure as much as
possible in order to have a steady state pressure as close to 50 bars as possible. The valve block was provided
with three ports, the inlet port connected to the outlet of the pump, the outlet port blocked with the help of a ballvalve and a reservoir port through which the leakage flow was allowed to flow to a beaker. A combination ofpressure sensors connected to a data acquisition system for monitoring of the pressure fluctuations in the inlet
port was employed as shown in figure 3.

Figure 3. Data Acquisition System for the inlet pressure and Electrical Connections for Solenoid Actuator Input

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RESULTS AND DISCUSSION
The leakage flow characteristics of the pressure reducing valve were determined by measuring the value of theflow rate that was obtained through the following simple equation:

Qleak Volume 100 ml

Time
Time

(3)

When Solenoid was ActuatedWhen the solenoid actuator was actuated the pressure at the valve’s outlet was recorded to be 37 bars. Since theoutlet was kept closed through a ball valve, the pressure of 37 bars was successfully achieved without any
noticeable fluctuations. The leakage value determined was 9ml/min.

When Solenoid was Turned off
In the absence of the solenoid’s actuation force, the outlet pressure was reduced to a value less than 0.4 bars at
which the leakage value was determined to be 12 ml/min.

Discussion
Table 1 shows a summary of the leakage-flow results.

Solenoid State
On
Off

Inlet Pressure (bars)
50
50

Table 1. Leakage flow Results Summary
Outlet Pressure (bars)
Flow Rate (ml/min)
37
9
0.4
12

The reason behind the increased flow rate when outlet pressure is increased can be explained by equation (1). At
high pressure the leakage is caused by the flow between the inner valve chamber and the reservoir port while at
low pressure the leakage is caused by the flow between the inlet, which is at 50 bars and the valve chamber
which is at 0.4 bars. Overall the average flow rate turned out to be 10.5 ml/min which for an inlet pressure of 50bars would result in a power loss of 0.875 watts.

CONCLUSION
The valve tested in this study had an average leakage of 10.5 ml/min. The leakage depends on the pressure. In
this study, the pressure at inlet was kept constant at 50bars and fluctuations in its value was minimized by theuse of an accumulator allowing us to test the valve for leakage under steady state. Power Loss was 0.875 watts.
In future, further details relating leakage with the geometrical construction and inlet pressure can be determined
and its effect on the power loss can be evaluated using a similar experimental setup as devised in this study.

ACKNOWLEDGMENT
This research was funded by Korea Institute of Machinery and Materials under the project title “Development of
the core equipment for liquid hydrogen supply systems”.

REFERENCES
1. SUTERA, S. P.; SKALAK, RICHARD, The History of Poiseuille's Law, Annual Review of Fluid Mechanics, 1993, 25:
1–19.
2. ISO 1219-1:2012, Fluid power systems and components — Graphical symbols and circuit diagrams — Part 1: Graphicalsymbols for conventional use and data-processing applications, 2012.

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Generation Mechanism of Flow Force Acting on Spool Valve
Fumio SHIMIZU* and Kazuhiro TANAKA***

Department of Physics and Information Technology, Kyushu Institute of Technology
680-4 Kawazu, Iizuka city, Fukuoka, 820-8502 Japan(E-mail: shimizu@phys.kyutech.ac.jp)**
Kyushu Institute of Technology

Abstract. In hydraulic systems, spool valves play an important role. When the spool in the valve moves to the left or
right, the internal flow path changes, and the hydraulic flow rate is controlled by the valve opening. Since high-pressure
fluid flows into the valve, it is necessary to predict the behavior of the spool with high accuracy. One of the problems
in spool valve is the hydraulic vibration caused by the high-pressure fluid. There is a need for a control valve that can
suppress the hydraulic vibration even if it is used alone. The purpose of this study is to visualize the flow field in thespool valve using three-dimensional computational analysis, and to clarify the generation mechanism of the flow force
acting on the spool. The jet flow around the spool generates high- and low-pressure regions, and the pressure difference
between the two regions generated the flow force.
Keywords: Spool valve, Flow force, Generation mechanism, Validation, Jet flow.

INTRODUCTION
Hydraulic systems are used in various fields, including transportation and construction machinery. The system
can be roughly classified as consisting of a tank, pump, piping, hydraulic control valve, and actuator. The
hydraulic control valve, such as a spool valve, plays an important role in a hydraulic system. A small spool isinstalled in the control valve. When the spool moves to the left or right, the flow path in the control valve changes,
and the flow rate of hydraulic oil is controlled by the valve opening. A small spool, about 1 [cm] in diameter,
controls the operation of a hydraulic circuit with several [MPa] of internal pressure. Therefore, the behavior of
the spool needs to be predicted with high accuracy [1]. One of the problems that spool valves have is the hydraulicvibration caused by the high-pressure fluid [2]. The most popular method of vibration damping is to combine
multiple control valves [3]. Recently, there is a need to reduce the overall size of the hydraulic machine, and astand-alone control valve that can suppress hydraulic vibration is required. Therefore, it is necessary to understandthe generation mechanism of the flow force acting on the spool, which is the cause of hydraulic vibration, and to
study the suppression method of the flow force. The purpose of the present study is to visualize the flow field ina spool valve using three-dimensional computational fluid dynamics (CFD) and to clarify the generation
mechanism of flow forces acting on the spool.

COMPUTATIONAL METHODSIn the present paper, numerical analysis was conducted using ANSYS CFX, that was thermal and fluid analysis
software [4]. CFX is a commercial software that is widely used around the world. The finite volume method isused as a discretization method, which provides excellent robustness and fluid conservations in numericalcalculations. The governing equations are shown below.

∇∙𝑈 =0(1)

𝜕𝑈
1+ ∇ ∙ (𝑈 × 𝑈) = − ∇𝑝 + 𝜇∇2 𝑈
𝜕𝑡
𝜌

(2)

Equation (1) is the continuity equation for the conservation of mass, and equation (2) is the Navier-Stokes
equations for the conservation of momentum.

RESEARCH TARGET AND COMPUTATIONAL CONDITIONSIn the present study, a two-port spool valve was chosen as the target of research. The approximate configurationof spool valve is shown in Figure 1. Figure 1(a) shows the structure and dimensions of the control valve. For the1

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The 11th JFPS International Symposium on Fluid Power 2020spool valve, one rectangular port was connected to the inlet side and another one to the outlet side. The crosssections of the inlet and outlet ports were square configuration, with a side length of 10 [mm] on the inlet and 8[mm] on the outlet. There was a spool with large and small diameter shafts, the diameter size was 16 and 8 [mm],respectively. The flow rate was controlled by the movement of the spool. It had a simple structure. Expressing thevalve opening on the inlet side as “S1” and the valve opening on the outlet side as “S2”, the following relationshipwas obtained,

S1 + S2 = 10 [mm].

(3)

There is an inlet throat (S1 ≤ 5.0 [mm]) that narrowed the flow path on the inlet side, and an outlet thraot (S2 ≤5.0 [mm]) that narrowed the flow path on the outlet side. In the present paper, numerical analysis was conductedfocusing on the inlet throat (S1).
Figure 1(b) shows the overhead view of the computational domain. The computational domain was created byenclosing area of the dashed lines in Figure 1(a). The diameter of the thick shaft of the spool was D=16 [mm],and this length was used as the reference length. The lengths of the inlet and outlet ports were set to 3D for both.
In numerical calculations, the spool was treated as having a fixed position. Five types of computational grids withdifferent valve openings S1 were created. The eccentricity and tilt of the spool were not considered. Each
computational grid have a total number of about 2 million grid points.

Figure 1(a). Spool Valve Configuration.
Figure 1(b). Computational Region.
Figure 1. Spool Valve Configuration and Computational Region.

Table 1 summarizes the computational conditions used in the numerical calculations. To explore the possibilityof the flow fluctuates with time, an unsteady analysis was performed. The time increments were set so that the
CFL condition would be about 10. The working fluid was oil, flowing at 20 [L/min]. The Shear Stress Transportmodel was used for the turbulence model. Numerical calculations were carried out for five different channel
geometries with different inlet throat S1.
Table 1. Computational Conditions.
Analysis typeTransient
Turbulence model

Shear stress transport

Total time

2.0  10−2 [s]Time step

5.0  10−5 [s]Density

850 [kg/m3]

Viscosity

3.5  10−5 [m2/s]

Inlet

20.0 [L/min]

OutletAverage static pressure 0.0 [Pa]Elements

2,000,000 [-]S1 length

0.2，0.5，1.0，2.0，5.0 [mm]2

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PREVIOUS RESEARCH
The flow force acting on the spool can be decomposed into an axial force acting in the direction of the spool axis
(Z-axis) and a lateral force acting in the radial (Y-axis) direction. Kondo et al. experimentally measured two typesof flow forces acting on a spool [5]. A three-dimensional numerical analysis was also performed using an original
numerical code. The flow visualization inside the valve and the flow force acting on the spool were investigated.
In the present paper, we compared the experimental and numerical results of Kondo et al. with the computational
results obtained by our calculations and confirmed the validity of the present computational method.

COMPUTATIONAL RESULTS AND DISCUSSIONS
Figure 2 shows the results of the present calculations compared with the experimental and numerical results ofKondo et al. Figure 2(a) shows the results of the axial force acting on the spool for five kinds of valve opening
S1. Figure 2(b) shows the results of the lateral force. The plotted marks in the figures were all time-averaged
values, and the error bars indicated the time variability range of the computational results. In Figure 2(a), the axial
force was almost zero for S1≥1.0 [mm]. However, as the inlet throat became narrow, the axial force increased
rapidly. The monotonic increase in axial force with decreasing inlet throat was quantitatively good agreementwith all three results. The error bars were expanded when the throat was small, so it suggested that time variability
occurred. For the lateral forces, present results were compared with the numerical results of Kondo et al. In Figure
2 (b), the value of lateral force was generally smaller than that of axial force. As the throat length decreases, thelateral force tended to increase and then decrease. When the throat length was narrow, unsteady fluctuations
appeared in the present results. A comparison of the time-averaged values showed good agreement with the resultsof Kondo et al., the present numerical analysis is considered to validity.

Figure 2(a). Axial Force.

Figure 2(b). Lateral Force.
Figure 2. Flow Force for Inlet Valve Opening S1.

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The 11th JFPS International Symposium on Fluid Power 2020

To investigate the generation mechanism of axial and lateral forces, the relationship between the flow patterninside the valve and the pressure distribution on the spool surface was examined. Figure 3 shows the instantaneous
streamlines and the pressure distribution on the spool surface. The case of S1=0.2 [mm] is shown. By observing
the streamlines, the flow inside the valve can be explained as follows.
(1) The hydraulic fluid that passed through the inlet throat flowed along the end face of the large-diameter shaftof the spool.
(2) After that, the hydraulic fluid collided with the small-diameter shaft on the inlet side.
(3) The hydraulic fluid traveled downstream along the thin shaft.
(4) The hydraulic fluid collided with the end face of the thick shaft on the outlet side.
(5) The hydraulic fluid flowed out to the outlet port with forming some vortex.
The figure on the right-hand side shows an enlarged view near the inlet throat region. The jet flowed in vigorously.
The outer circumference parts of the thick shaft were at the root of the jet, which were a low-pressure region due
to the fast velocity of the jet. On the thin shaft, high-pressure region appeared because of the collision of the jet.
The pressure difference between low- and high-pressure regions generated the flow force.

Figure 3. Streamlines and Pressure Distribution for S1=0.2 [mm].

To clarify the mechanism of axial force generation, the flow field inside the valve and the pressure distribution
acting on the end faces of the thick shaft were investigated. The results for S1=0.5 [mm] are shown in Figure 4.
The velocity vectors in the YZ section and the pressure distribution at the two end faces of the thick shaft are
shown. On the left side surface, a low-pressure region was extended by the jet blow. In particular, the pressure
was extremely low at the outer circumference part, which was just fast jet area. On the other hand, the right-side
end face of the thick shaft was a high-pressure region because the flow along the thin shaft collided with the end
face. It was found that the axial force was generated by the pressure difference acting on the left and right end
faces.

Figure 4(a). Velocity Vectors Plots.
Figure 4(b). Pressure Distribution on End Faces.
Figure 4. Velocity Vectors Plots and Pressure Distribution for S1=0.5 [mm].

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The time variations of the flow field were investigated for S1=0.2 and S1=0.5 [mm]. The velocity distributions atthree different times in the YZ section are shown in Figure 5. The high-pressure area on the spool surface wasindicated with solid triangle marks, and the low-pressure area was indicated with blank triangle marks. First, we
focused on S1 = 0.2 [mm]. The hydraulic fluid that flowed in from the inlet throat became a jet flow and flowed
along the end face of the thick shaft and the thin shaft body. The flow formed some vortex near the end face ofthe thick shaft on the outlet side. Although the jet shape at the upper side remained almost unchanged with time,
the jet shape at the lower side changed its direction significantly with time. The lower jet collided with the thin
shaft or not collided the shaft. Next, we focused on the pressure distribution acting on the spool. The end face ofthe thick shaft showed a low-pressure region on the inlet side (indicated by the blank marks), while the outlet side
showed a high-pressure region. As a result, there was the pressure difference acting on the left and right surfaces,
which acted on an axial force in the +Z (right) direction. The axial force did not change much over time. On theother hand, the pressure on the surface of the small-diameter shaft showed time variation depending on thepresence or absence of the jet impact. When the vortex formed on the outlet side collided with the thin shaft, it
was observed that a high-pressure region was also formed on the outlet side. The pressure distribution on the thin
shaft surface changed significantly with time. Therefore, it was confirmed that unsteady changes in the lateralforce occurred.
When S1 = 0.5 mm, the obliquely oriented jet impacted the thin shaft. It was also observed that the flow along thethin shaft changed the direction with time. However, the change was slight, indicating that the unsteady variation
was reduced.

Figure 5. Switching flow at S1=0.2 and 0.5 [mm].

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Since the spool position is fixed in the present study, the spool vibration cannot be reproduced. However, spool
vibration may occur in actual spool valves due to the action of unsteady flow forces. To investigate the state ofspool vibration using CFD, it is necessary to perform fluid/structure coupled analysis that takes spool movement
into account.
There is a small gap of several micrometers between the spool and the sleeve. The difference in the gap may causea change in the flow force. This small gap is not considered at present, but we would like to conduct a numericalanalysis taking the gap into account in the future work.

CONCLUDING REMARKS
A two-ports spool valve was numerically investigated using three-dimensional computational fluid dynamics.
Flow forces acting on the spool in two directions were investigated. As the results of investigating the magnitudeof the flow forces against the change in inlet valve opening, the results of the present calculations were in good
agreement with the results of the previous study. The present calculations were confirmed to be valid. It was foundthat the flow force acting on the spool was caused by the pressure difference between the high-pressure region
generated by the impact of the jet and the low-pressure region through which the high-speed jet passed. Fluctuationof the jet flow occurred when the valve opening became narrow. The high-pressure region changed with time dueto the time variation in the jet flow direction. It was found that the unsteady fluctuation of the jet flow had asignificant effect on the unsteady variation of the flow forces.

ACKNOLWLEDGMENTS
We would like to express our gratitude to Mr. Saiki Tanaka for his great cooperation in the present research.

REFERENCES
1. LEE, S. Y. and BLACKBURN, J. F., Steady-State Axial Forces on Control-Valve Pistons, Transactions of the American
Society of Mechanical Engineers, 76, 1952, pp. 1005-1011.
2. KUMAGAYA, H. and MASAKI, K., Your First Hydraulic System (in Japanese), Gijutsu-Hyohron Co., Ltd., 2009, pp.
100-153.
3. ICHIKAWA, T. and HIBI, A., Oil-Hydraulic Engineering (in Japanese), Asakura Publishing, 1979, pp. 88-118.
4. ANSYS Inc, ANSYS CFX-Solver Modeling Guide Release 13.0, 2010, pp. 13-15, pp. 149-152.
5. KONDOH, Y., et al., Analysis of Flow Forces Acting on a Spool Valve (1st Report, Non-Uniformity of Flow Pattern andMomentum Flux in the Azimuthal Direction) (in Japanese), Transactions of the Japan Society of Mechanical Engineers,
Series B, 65-639, 1999, pp. 3577-3585.

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The Effect of the Spline Coupling on the Rotating Assembly
Tilt Behavior in a High-speed Axial Piston Pump
Haogong XU*, Junhui ZHANG*, Weidi HUANG*, Bing XU*, Fei LYV*, Xiaochen HUANG*,*

State Key Laboratory of Fluid Power and Mechatronic Systems, Zhejiang University,
38 Zheda Road, Xihu, Hangzhou, 030027, China(E-mail: xuhaogong@zju.edu.cn)

Abstract. The rotating speed of the axial piston pump is increasing to meet the requirement of a higher power-toweight ratio. However, the cylinder block in the pump tilts severely at a high rotating speed, which seriously affects
the reliability and efficiency of the pump. The cylinder block is the main component of the rotating assembly, so thetilt behavior of the cylinder block is analyzed through a proposed dynamic model of the rotating assembly in thepaper. The elastic deformation of the shaft, the nonlinear bearing characteristics of the oil film, and the reaction of
spline coupling are taken into account. Then the tilt angle of the cylinder block and the meshing status of the spline
coupling are analyzed. Finally, the influence of the working conditions is investigated to offer an insight into the tilt
behavior of the rotating assembly in a high-speed axial piston pump.
Keywords: Rotating assembly, cylinder block, transient behavior, finite element model, spline coupling

INTRODUCTION
The high-speed axial piston pump is widely used in the aerospace industry and the defense industry due tothe high efficiency and high power-to-weight ratio. To meet the demand for a higher power-to-weight, therotating speed and the outlet pressure of the pump are increasing gradually. However, the leakage of the pumpincreased dramatically at a high rotating speed which reduces the efficiency of the pump. The leakage mainly
comes from the three friction pairs: the swash plate/slipper pair, the cylinder block/piston pair, and the valve
plate/cylinder block pair. The three friction pairs are shown in Figure 1. The cylinder block is driven by theshaft to slide on the valve plate surface. The pistons are forced to rotate around the shaft axis and reciprocate
within the cylinder bores at the same time. The slippers slide on the inclined swash plate. The oil film within thethree friction pairs plays an important role in the lubrication but generates the leakage which decreases the pumpefficiency.

Figure 1. The Inner Structure of the High-speed Axial Piston Pump

The leakage within the friction pairs depends on the pressure and the gap sizes of the friction pairs. In thehigh-speed pump, the valve plate/cylinder block pair has a larger scale than the other two friction pairs, and it
contributes more than 70% of the total leakage of the pump [1]. On the other hand, the oil film thicknesses ofthe swash plate/slipper pair are limited mechanically by the hold-down mechanism (Retainer cover), while thecylinder bores restrict the oil film thicknesses of the cylinder block/piston pairs. However, the cylinder block ismainly pressed towards the fixed valve plate by the pressurized oil and the central spring, and the maximum oilthickness of the valve plate/cylinder block pair is barely restricted. As a result, the cylinder block is more likely
to tilt under high pressure and high rotating speed conditions, which has a noticeable effect on the pumpefficiency.

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The tilt behavior of the cylinder block determines the oil film shape of the valve plate/cylinder block pair.
Many researchers have conducted the thickness distribution measurements of the valve plate/cylinder block pairin industrial axial piston pumps under the actual operating conditions [2–6]. And all the results showed that
remarkable tilting of the cylinder block has occurred, and the operating conditions, especially the outlet pressure,affected the tilt behavior of the cylinder block obviously. On the other hand, theoretical analyses have been
carried out to investigate the bearing characteristics of the valve plate/cylinder block pair. Many types of themodels, such as the analytical model [7,8], elastohydrodynamic model [9–11], and the thermal
elastohydrodynamic model [1,12–15], have been established. The oil film thickness, temperature distribution,and the deformations of the sliding surfaces in the valve plate/cylinder block pair were analyzed based on thesemodels. However, these lubrication models mainly focused on the nonlinear bearing characteristics of the oilfilm but neglected the reactions from other parts in the pump.
Moreover, several researchers paid attention to the interactions between the moving parts in the axial piston
pump/motor based on the multi-body dynamic models. The first type of dynamic model was established under
the commercial software environment [16–18], and the detailed geometry could be taken into account. The
second type was the lumped parameter analytical model [19–21], and the whole pump was modeled as a massspring-damper system. However, the movements of the moving parts were over-constrained, and the bearing
characteristics of the oil film were always linearized, which would bring obvious errors under a wide operating
range.
The cylinder block movement is affected by the connected parts, such as the piston-slipper assemblies, theshaft, and the valve plate. The previous studies mainly focused on the industrial axial piston pump/motor. But ina high-speed axial piston pump, the nonlinear speed-dependent bearing characteristics of the oil film are much
more prominent. And the radial force due to the inclined swash plate results in a slight inclination and bendingof the shaft, which may affect the meshing status of the spline coupling. However, the underlying interactions
between the relative moving parts are still not completely understood. In this paper, a dynamic model of therotating assembly is proposed to investigate the tilt behavior of the cylinder block in a high-speed axial piston
pump. The rest of this paper is arranged as follows: the detailed dynamic model is presented in Section 2; thetheoretical results are presented and discussed in Section 3; finally, the conclusion and outlook are drawn in
Section 4.

DYNAMIC MODEL OF THE ROTATING ASSEMBLY
The rotating assembly of the high-speed pump is mainly comprised of the shaft, the cylinder block, and 2
bearings. The dynamic model of the rotating assembly is shown in Figure 2. The finite element method isadopted to model the shaft and the cylinder block, considering the elastic deformation. A theoretical model isestablished to investigate the reactions of the spline coupling between the shaft and the cylinder block. The
nonlinear bearing characteristics of the oil film are taken into account based on the one-dimensional Reynolds
equation.

Figure 2. The Finite Element Model of the Rotating Assembly

Finite Element Model of the Rotating Assembly
The finite element model consists of the shaft and the cylinder block. The shaft and the cylinder block aremodeled as Timoshenko beam considering their elasticity and inertia. As shown in Figure 2, the shaft is divided
into 20 elements by 21 Nodes, in which the spline on the shaft consists of 11 elements. Due to the axisymmetriccross-section of the rotating assembly, the torsional displacement is uncoupled with lateral displacements. The
axial movement of the shaft is limited by the preloaded angular contact ball bearing. As a result, each shaft

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element contains 2 translation degrees of freedom(u,v) and 3 rotary degrees of freedom(θx,θy, θz). According to
the non-conservative Lagrange equation, the motion equations of the shaft can be expressed as:   M s  q s    G s  q s    K s  q s   F s 
 
 

(1)

where superscript ‘s’ denotes the shaft unit, M, G, K are the mass matrix, gyroscopic matrix, and stiffness
matrix respectively, Ω is rotating speed, q is the generalized coordinates vector, F denotes generalized force
vector exerted on shaft unit.
The contact status of the spline on the cylinder block also should be taken into account, and the cylinder
block is divided into 11 elements by 12 nodes. The additional axial movement of each node in the cylinder
block is considered compared with the shaft node, since the axial movement of the cylinder block w has aneffect on the oil film within the cylinder block/valve plate pair. Therefore, the motion equations of the cylinder
block can be obtained   M c  q c    G c  q c    K c  q c   F c 
 
 

(2)

where superscript ‘c’ denotes the cylinder block unit. Finally, the motion equations of the rotating assembly
can be expressed as

M s



 
 q s G s 
Mc cq   q s   K s  Gc   c  q  q s   F s     
K c  q c   F c (3)

The Bearing Capacity of the Valve Plate/Cylinder Block Pair
The oil film within the valve plate/cylinder block pair is illustrated in Figure 3. According to the geometry,the oil film can be divided into the inner sealing land region, the main kidney land region, and the outer sealing
land region. The main kidney land region consists of the pressure ports and the transitional areas of the valve
plate, and the pressure in the transitional areas varies periodically as the kidney port of the cylinder block entersand leaves the transitional areas. The pressure in the main kidney land region mainly depends on the pressure in
the piston chamber and the pressure port of the valve plate, and it is not affected by the oil film shape.

  3 p h rh
  6  r
r 
r 

Figure 3. The Oil Film Within the Valve Plate/Cylinder Block Pair

On the contrary, the oil film shape has a significant influence on the pressure distribution in the sealing lands.
The two-dimensional Reynolds equation is usually adopted to model the bearing capacity of the oil film, but it
can only be solved numerically by the finite element method(FEM) or the finite volume method(FVM), whichslows down the calculation efficiency. Fortunately, the width of sealing lands is small enough compared withthe radius in the valve plate/cylinder block pair so that the circumferential pressure gradient can be neglected.
Then oil film pressure distribution in each sealing land can be obtained based on the one-dimensional Reynolds
equation(4)
where r is the radius of the oil film, h is the oil film thickness, p is the pressure distribution of the oil film, ω isthe rotating speed of the cylinder block, μ is the oil viscosity. When neglecting the elastic deformation, the oilfilm thickness distribution is expressed as(5)h  h 0  rm cos 

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where α is the tilt angle of the cylinder block, rm is the average radius of the sealing land, the tilt angle α and thecentral gap height h0 are related to the movement of the cylinder block. Then the pressure distribution in eachsealing land is obtained2
 ln  rin / r   routln  rin / r   rin2 ln  rout / r  2 
 ln( r / rout ) 3 rm sin 
psl  r ,  R   pin 

r 
 
  pout 3 ln  rin / rout  ln( rin / rout )  ln  rin / rout   2  h0   rm cos   (6)where the subscript ‘in’ and ‘out’ are inner pressure boundary and outer pressure boundary of the sealing land,
respectively.
Besides, the asperity contact may occur when the oil film thickness is rather small. The asperity contactpressure can be obtained by the Greenwood –Tripp model [22]h
pc  K E F5/ 2  
 

(7)

where K' is the elastic factor, E' is the composite elastic modulus, F5/2(h/σ) is the form function, and σ is the
combined standard deviation of the asperities.
Finally, the bearing force and moments of the oil film are given by integrating all the pressure components
numerically
Fvz    prdrd2
 M vx  -   pr cos drd2
 M yx    pr sin  drd(8)

The Bearing Capacity of the Spline Coupling
In the high-speed axial piston pump, the shaft and cylinder block are connected by an involute spline
coupling, which not only transfers the driven torque but also supports the cylinder block in the radial direction.
The radial force exerted on the cylinder block due to the inclined swash plate results in the deformation of the
shaft and the cylinder block, and it changes the meshing status of the spline coupling. Therefore, the bearing
capacity of the spline coupling is investigated through FEM, and the splines on the shaft and cylinder block are
divided into 11 shaft elements, as shown in Figure 2. In each spline coupling element, only the parallel
misalignment is taken into account, as shown in Figure 4.

Figure 4. The Misalignment of the Spline Coupling

The meshing status of the spline coupling element is illustrated in Figure 5. When the spline coupling iscompletely aligned, the clearances between all the teeth of the spline are evenly distributed

 initial 1
 cos 2

(9)where Δ is the backlash of the spline coupling, and γ is the pressure angle. However, the relative movement of
the cylinder block and shaft changes the clearance distribution. The relative movement for each spline couplingelement consists of the radial movement and torsional movement, and the additional clearance due to the two
types of movements can be obtained

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rij   si cos(2j      si );n2ti   rs  si   ci 

( j  1, 2,...n)

(10)
(11)

where σsi, ξsi are the radial movement and the azimuth angle of the relative movement for the i spline couplingelement, respectively. n is the teeth number, rs is the pitch diameter of the spline. φsi and φci are the torsional
movements for the ith spline coupling element on the shaft and the cylinder block, respectively. Finally, theclearance distribution is expressed as ij   initial  rij  ti(12)th

Figure 5. The Meshing Status of the Single Spline Coupling Element

The meshing forces and the relevant force and torque on the shaft can be obtained  K sp L ij ,  ij  0 
Fij  0,  ij  0 


 2
Fxsi   Fij cos j   
2
 n

 2
Fysi   Fij sin j   2 nTzsi    Fij r cos

(13)

(14)

(15)
(16)

where Ksp is the spline tooth constant stiffness, L is the axial length of the spline coupling element.

SIMULATION RESULTS AND DISCUSSION
The Tilt Behavior of the Rotating Assembly
The rotating assembly is mainly subjected to the radial forces and moments from the piston-slipper
assemblies and the axial force from the pressurized oil in the piston chambers [23]. The tilt behavior of therotating assembly under these exerted forces and moments is investigated in a high-speed axial piston pumpwith 9 pistons.
The tilt behavior of the cylinder block (the bottom surface) under 9000rpm and 6MPa is shown in Figure 6.
The tilt angle and azimuth angle of the cylinder block fluctuate periodically, and the main wave contains 9 peaks
during a cycle, which is consistent with the number of pistons. The azimuth angle varies between 127 and 131°,and it means the cylinder block tilts towards the area near the outer dead center on the high-pressure port of thevalve plate.

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10-4

131

Tilt azimuth angle( °)Tilt angle(rad)

2.8
2.6
2.4
2.2

130
129
128
127

2
0

200

400

600

800

0

1000

200

400

600

800

1000

Rotating angle(°)

Rotating angle(°)a)b)

Figure 6. The Tilt Behavior of the Cylinder Block: a).The Tilt Angle; b). The Tilt Azimuth Angle

The tilt behavior of the rotating assembly is obtained based on the average displacement of all the Nodes, as
shown in Figure 7. It can be seen that the shaft has bending deformation along the x and y axes, and thedeformation along the y-axis is significantly greater than that along the x-axis. The incline of the shaft also canbe found under the external load due to the elastic bearings. And the shaft inclines upward along the x-axis andinclines downward along the y-axis. The axes of the shaft and the cylinder block spline are obviously out ofalignment. The movements of the spline coupling axes are extracted as shown in Figures 7c and 7d. The
bending deformation and incline of the cylinder block axis can be found. The relative deflection between the
spline centerlines of the cylinder block and shaft is much larger along the y-axis than that along the x-axis. And
the left side of the spline coupling shows the more significant relative movement along the y-axis. The clearance
distribution of the spline teeth is shown in Figure 7e. The negative values indicate that all the teeth are engaged
effectively. The spline tooth pairs at node 1 present larger deformations(smaller clearances), and more nonuniform deformation distribution can be found at node 1. It results from not only a larger relative radial
movement at the left side of the spline coupling but the largest relative torsion angle at node 1 since it is close to
the input torque of the shaft. The maximum deformation of the tooth pair reaches 8.5μm and is located at the
tooth pair near the negative x-axis at node 1(the tooth number starts from the positive x-axis and is numbered
anticlockwise). With the node number increasing, the maximum deformation gradually decreases to about 3μm,
while the location moves to the positive x-axis.

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Figure 7. The Tilt Behavior of the Rotating Assembly: a). The Tilt Behavior of the Rotating Assembly along the x-axis;b). The Tilt Behavior of the Rotating Assembly along the y-axis; c). The Movement of the Spline Coupling Axes along the xaxis; d). The Movement of the Spline Coupling Axes along the y-axis; e). The Clearance Distribution of the Spline Teeth

The Influence of the Working Condition on the Tilt Behavior of the Rotating Assembly
The influence of working conditions on the tilt behavior of the cylinder block is shown in Figure 8. It can beseen that the higher the speed is, the greater the average tilt angle of the cylinder block is. The influence of therotating speed on the tilt angle is much more obvious under a lower outlet pressure. There is a speed thresholdof about 6000 rpm. When the speed is lower than the threshold speed, the pressure increase will significantly
enhance the tilt angle of the cylinder block. And the effect of pressure on the tilt angle is opposite when the
speed is higher than the threshold speed. In addition, when the outlet pressure increases, the tilt azimuth angle
also increases, and it means that the location of the minimum oil film thickness moves to the center of the highpressure port of the valve plate. And the tilt azimuth angle first decreases then increases with the increasing
rotating speed. The minimum oil film thickness of the cylinder block/valve plate pair decreases with the increaseof pressure, which indicates that the asperity contact of the valve pair is more severe under high pressure.
On the other hand, the minimum oil film thicknesses under different pressure conditions decrease first andthen increase with the rotation speed, as shown in Figure 8c. The main reason is that with the increase ofrotation speed, the centrifugal forces of the piston-slipper assemblies increase rapidly, and the resulting
increased tilt moment intensifies the tilt of the cylinder block. With the further increase of the speed, thehydrodynamic effect of the oil film is enhanced rapidly, which leads to the overall increase of the oil film
thickness, and it counteracts the decrease of minimum oil film thickness caused by overturning. Under low
outlet pressure, the minimum oil film thickness is larger. It is noticed that the minimum oil film thickness
increases sharply at a high rotating speed, especially under the low outlet pressure. At this time, the leakage ofthe pump will increase sharply, which seriously affects the efficiency of the pump.

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10-4

160

Tilt azimuth angle( °)Tilt angle(rad)

4

3MPa
6MPa
12MPa
18MPa

3
2
1

3MPa
6MPa
12MPa
18MPa

150
140
130
120
110

0
3000 4000 5000 6000 7000 8000 9000 10000

100
3000 4000 5000 6000 7000 8000 9000 10000

Rotating speed(rpm)Rotating speed(rpm)b)
Minimum oil film thickness(mm)

a)
3.5

10-3
3MPa
6MPa
12MPa
18MPa

3
2.5
2
1.5
1

0.5
3000 4000 5000 6000 7000 8000 9000 10000

Rotating speed(rpm)c)
Figure 8. The Tilt Behavior of the Cylinder Block under Different Working Conditions: a). The Tilt Angle;b). The Tilt Azimuth Angle; c). The Average Minimum Thickness of the Oil Film

The effect of the outlet pressure on the spline coupling movement is shown in Figure 9. It can be seen thatthe pressure has little influence on the movements of the spline coupling axes along the x-axis. With the increaseof pressure, the displacements of all the Nodes on the shaft and cylinder block spline axes decrease, but thechange of the relative movement between the axes is not significant. On the other hand, the outlet pressure
dramatically affects the spline coupling movement along the y-axis. With the increase of pressure, the increasingload on the cylinder block is transferred to the shaft, so the axes movements of both the shaft and cylinder blockincrease rapidly. Higher outlet pressure leads to a more considerable bending deformation of the shaft, and therelative displacement of the spline axes on the shaft and cylinder block increases obviously, especially the left
side Node. The increased relative displacement at the left side node always results in the larger deformation of
spline tooth as shown in Figure 9c and 9d. It indicates the increasing meshing force of the tooth pair, which
provides a higher anti-tilt moment for the cylinder block. The axes of the shaft and the cylinder block spline
gradually separated from the intersection with the increasing outlet pressure, and the spline axis of the cylinder
block is entirely above the shaft spline axis under 18MPa. In addition, the driven torque also increases when theoutlet pressure rises, and it results in the larger relative twist angle between the shaft and the cylinder block.
That’s why the deformations, as well as the their distribution, of the spline teeth under high pressure are muchlarger than that under low pressure.

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a)

b)
m

m
-1
-2

10

-3

5

-4

2

4

6

8

15

Spline tooth number

Spline tooth number

15

-10

10
-15

5
-20

2

10

Spline node on the cylinder block

4

6

8

10

Spline node on the cylinder blockc)d)Figure 9. The Effect of the Outlet Pressure on the Spline Coupling Movement: a). The Movement along the x-axis;b). The Movement along the y-axis c). The Clearance Distribution of the Spline Teeth under 3MPa;d). The Clearance Distribution of the Spline Teeth under 18MPa;

The effect of the rotating speed on the spline coupling movement is shown in Figure 10. It can be found thatthe rotation speed has a noticeable influence on the movements of the spline coupling axes, especially along they-axis. The inclinations along the x-axis of the spline coupling axes increase with rotating speed, and the relative
inclination between the axes also increases synchronously. In the y-axis, with the increase of the rotation speed,
the tilt degree of the cylinder block is intensified due to the increased inertia tilt moment of the piston-slipper
assemblies and the spline axis of the cylinder block inclines downward. The spline axis of the shaft also inclines
downward. At the same time, the axes of the shaft and the cylinder block spline gradually intersect with eachother as the rotating speed increases. Comparing Figure 10c and 10d, we can find that the maximum
deformation of the spline tooth and the range of the deformation increase slightly with the rotating speed. The
location of the maximum deformation on the right side of the spline coupling changes from the negative x-axis
to the positive x-axis when the rotating speed increases to 10000rpm.
The spline coupling mainly balances the radial force along the y-axis and the tilt moment around the x-axis.
When the outlet pressure increases, the radial force increases more significantly than the tilt moment. At thistime, all the spline coupling Nodes generate the bearing forces along the negative y-axis on the shaft so as tobalance the increasing radial force under high pressure. But increasing rotating speed enlarges the tilt moment

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a)

b)
m

m
-3
-4

10

-5

5

-6

Spline tooth number

15

Spline tooth number

-2

15

-3
-4

10

-5
-6

5

-7
-8

-7

2

4

6

8

2

10

Spline node on the cylinder block

4

6

8

10

Spline node on the cylinder blockc)d)Figure 10. The Effect of the Rotating Speed on the Spline Coupling Movement: a). The Movement along the x-axis;b). The Movement along the y-axis; c). The Clearance Distribution of the Spline Teeth under 3000rpm;d). The Clearance Distribution of the Spline Teeth under 10000rpm

while barely has an effect on the radial force. As a result, the Nodes of the spline coupling at the left and rightsides provide the opposite bearing force to balance the inertia tilt moment at the high rotating speed. On theother hand, the maximum tilt angle of the cylinder block appears under the condition of low outlet pressure andhigh rotating speed. It means that the ability to balance the tilt moment of spline coupling needs to be further
improved to reduce the maximum tilt angle of the cylinder block. Increasing the contact length of spline
coupling, or reducing the elastic deformation of the shaft by optimizing the shaft geometry, might be theeffective method, which needs to be investigated in the following research.

CONCLUSION
In this paper, a dynamic model of the rotating assembly in a high-speed axial piston pump is presented. Inorder to investigate the bearing characteristic of the spline coupling, a contact model of the meshing teeth
considering the elastic deformation is implemented. Then the dynamic model is adopted to investigate thedynamic behavior of the cylinder block and the rotating assembly.
The simulation results show that the cylinder block mainly tilts towards the area near the outer dead centeron the high-pressure port of the valve plate. The elastic deformation and inclination of the shaft can be observed
obviously, especially along the y-axis. And relative deflection between the spline centerlines of the cylinder
block and shaft results in the uniform clearance distribution of the spline coupling, which generates the reaction
forces to balance the radial force and tilt moment from the piston-slipper assemblies, and the maximum
deformation of the tooth pair is found in the left side of the spline coupling.
The working condition has a significant influence on the tilt behavior of the rotating assembly. The tilt angleof the cylinder block increases with the increasing rotating speed, especially under low outlet pressure. With theincrease of pressure, a more considerable bending deformation of the shaft is observed, and the relative
displacement of the spline axes on the shaft and cylinder block increases obviously. The axes of the shaft andthe cylinder block spline gradually separated from the intersection to balance the increasing radial force underhigh pressure. With the increase of the rotation speed, the relative displacement of the spline axes on the shaft
and cylinder block also increases. And the axes of the shaft and the cylinder block spline gradually intersectwith each other as the rotating speed increases to balance the inertia tilt moment at the high rotating speed. The

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spline coupling node at the left side always has the maximum relative movement, and it is affected by thepressure and rotating speed obviously.
In future work, more detailed models for the rotating assembly will be included, such as the bearing models.
And the effect of the geometry parameters of the shaft and the cylinder block will be investigated to help carryout the optimization of the rotating assembly.

ACKNOWLEDGMENTS
The authors gratefully express their appreciation for the financial supports from the National Key R&D
Program of China (Grant No. 2019YFB2005101), National Outstanding Youth Science Foundation of China(Grant No. 51922093), and the National Natural Science Foundation of China (Grant No. 51890882), theNatural Science Foundation of Zhejiang Province(Grant No. LQ21E050022).

REFERENCES

1.

Han, L., Wang, S., Zhang, C., A partial lubrication model between valve plate and cylinder block in axial piston pumps,
Proc. Inst. Mech. Eng. Part C-J. Eng. Mech. Eng. Sci. 229 (2015), 3201–3217.
2. Kim, J.-K.,Jung, J.-Y., Measurement of fluid film thickness on the valve plate in oil hydraulic axial piston pumps (I)bearing pad effects-, KSME Int. J. 17 (2003), 246–253.
3. Xu, H., Zhang, J., Sun, G., et al., The direct measurement of the cylinder block dynamic characteristics based on a noncontact method in an axial piston pump, Measurement 167 (2021):108279.
4. A. Shorbagy, R. Ivantysyn, J. Weber, An experimental approach to simultaneously measure the temperature field andfluid film thickness in the cylinder block/valve plate gap of an axial piston pump, THMT-18. Turbulence Heat andMass Transfer 9 Proceedings of the Ninth International Symposium On Turbulence Heat and Mass Transfer., Begel
House Inc.,2018.
5. Wegner, S., Gels, S., Jang, D.S., et al., Experimental Investigation of the Cylinder Block Movement in an Axial Piston
Machine, Proceedings of the ASME/BATH 2015 Symposium on Fluid Power and Motion Control, Chicago, USA,2015.
6. Deeken, M., Simulation of the Tribological Contacts in an Axial Piston Machine, ASME International MechanicalEngineering Congress and Exposition. 2004, 47101: 71-75.
7. Manring, N.D., Tipping the cylinder block of an axial-piston swash-plate type hydrostatic machine, J. Dyn. Syst.,
Meas., Control 122.1 (2000), 216–221.
8. J. M. Bergada, J. Watton, S. Kumar, Pressure, Flow, Force, and Torque Between the Barrel and Port Plate in an Axial
Piston Pump, J. Dyn. Syst., Meas., Control 130.1 (2008).
9. Ivantysynova, M.,Baker, J., Power Loss in the Lubricating Gap between Cylinder Block and Valve Plate of Swash
Plate Type Axial Piston Machines, International Journal of Fluid Power 10 (2009), 29–43.
10. C HUANG,Ivantysynova, M., A New Approach to Predict the Load Carrying Ability of the Gap between Valve Plate
and Cylinder Block, Bath Workshop of Power Transmission and Motion Control PTMC 2003, Bath, UK, 2003.
11. Jihai, J.,Weipeng, Y., An Approach to Predict Wear Distribution of Valve Plate in Elasto-Hydrodynamic Lubrication,
IEEE Access 7 (2019), 86789–86797.
12. A. Schenk, M. Zecchi, M. Ivantysynova, Accurate Prediction of Axial Piston Machine's Performance Through a
Thermo-Elasto-Hydrodynamic Simulation Model, Proceedings of the ASME/BATH 2013 Symposium on Fluid Power& Motion Control, Sarasota, USA, 2013.
13. CHACON, R.,Ivantysynova, M., An Investigation of the Impact of Micro Surface on the Cylinder Block/Valve Plate
Interface Performance, Proceedings of the 8th FPNI Ph.D Symposium on Fluid Power, Lappeenranta, Finland,2014.
14. M. Zecchi, A Novel Fluid Structure Interaction and Thermal Model to Predict the Cylinder Block / Valve Plate
Interface Performance in Swash Plate Type Axial Piston Machines, Diss. Purdue University, 2013.
15. Ernst, M.,Vacca, A., Hydrostatic vs. hydrodynamic components of fluid pressure in the tribological interfaces of axial
piston machines, Tribology International 157 (2021), 106878.
16. Roccatello, A., Manco, S., Nervegna, N., Modelling a Variable Displacement Axial Piston Pump in a Multibody
Simulation Environment, J. Dyn. Syst., Meas., Control 129 (2007), 456-468.
17. ZHANG, B., Xu, B., XIA, C., et al., Modeling and Simulation on Axial Piston Pump Based on Virtual Prototype
Technology, CJME 22 (2009), 84–90.
18. Milind, T.R.,Mitra, M., A Study on the Dynamics and Vibration Behavior of an Axial Piston Pump Using Combined
MBD/FE Approach, Procedia Engineering 144 (2016), 452–460.
19. Chen, H.X., Chua, P.S.K., Lim, G.H., Dynamic vibration analysis of a swash-plate type water hydraulic motor, Mech.
Mach. Theory 41 (2006), 487–504.
20. Ye, S., Zhang, J., Xu, B., et al., A theoretical dynamic model to study the vibration response characteristics of an axial
piston pump, Mech. Syst. Signal Proc. 150 (2021):107237.

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21. Huang, H., Wu, G., Wu, Y., et al., Identification of acoustic sources for bent-axis axial piston motor under variableloads, J Sound Vib 468 (2020), 115063.
22. N. Patir, Effects of Surface Roughness on Partial Film Lubrication Using an Average Flow Model Based on Numerical
Simulation, Ph. D. Thesis, Northwestern University, 1978.
23. Zhang, J., Chao, Q., Xu, B., Analysis of the cylinder block tilting inertia moment and its effect on the performance ofhigh-speed electro-hydrostatic actuator pumps of aircraft, Chin. J. Aeronaut. 31 (2018), 169–177.

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First-order Trajectory Sensitivity Analysis of Multi-level Pressure
Switching Control System
Jing YAO * ** ***, Pei WANG *, Xinhao LI*, Yuwang CHENG* and Xiaoming CAO*****

School of Mechanical Engineering, Yanshan University
438, Qinhuangdao, Hebei, China(E-mail: wpei@stumail.ysu.edu.cn)**
Key Laboratory of Advanced Forging & Stamping Technology and Science of Ministry of Education ofChina, 438, Qinhuangdao, Hebei, China***
Hebei Key Laboratory of Heavy Machinery Fluid Power Transmission and Control, Yanshan University
438, Qinhuangdao, Hebei, China****
CCTEG Coal Mining Research InstituteBeijing, China
Abstract. Some variables have some uncertain influence on the control accuracy of the Multi-level Pressure
Switching Control System (MPSCS), but the importance and degree are not known clearly, which adds difficulty to
the targeted compensation control for high precision. This paper utilizes sensitivity analysis to explore the output
model of MPSCS response to varying fifteen variables and quantifies those variations for control compensation andoptimization. The first-order trajectory model of sensitivity based on the state equation of the MPSCS was built toanalyze the sensitivity-time history curves of each variable. The indicators of peak sensitivity and mean sensitivity
are proposed to quantitatively evaluate the effects of interactions among variables. The simulation and experiments
test the influence of model input variables and identify the variable that has the most influence on output
characteristics by indicators. Hence, the results serve as a great basis for control optimization and compensation of
the MPSCS.
Keywords: Sensitivity analysis, Multi-level pressure, Digital Hydraulic, Quantitatively analysis

INTRODUCTION
Compared with the traditional valve-controlled technique, digital hydraulics is a cutting-edge alternative of ﬂuid
power due to the advantages of low cost, high efﬁciency, control ﬂexibility, and robustness [1-3]. It is composedof several on/off valves actively controlling system outputs [4-5]. Besides, it covers a wide range of applications
including digital hydraulic valves, pumps, actuators, and switched hydraulics, as well as digital hydraulic power
management systems, which all promise high energy efficiency. Switched hydraulics is a sub-domain of digital
hydraulics, including the switching valve-controlled system and the switching hydraulic power supply system.
The switching valve-controlled system [6-8], such as the ABS brakes of cars, switching converters, SIHS, etc.,
is used to adjust or control flow and pressure [9]. The switching hydraulic power supply system [10-12] uses acontroller to improve energy efficiency depending on the current load by switch on/off valves. The group at
Aalborg University investigated discrete fluid power force systems for wave energy converters (WEC) andanalyzed the system configuration [13]. The discrete displacement cylinder (DDC) with several working
chambers connected with many pressure supply levels based on Force Switching Algorithm (FSA) to reducechattering and energy loss [14-16]. The concept of STEAM is introduced in [17], which develops differentpressure rails across the on/off valves to match loads and pumps in real-time. The research group at Yanshan
University [18-19] applied the two-level pressure to a forging hydraulic press and utilize on/off valves toachieve multi-level pressure independent metering in hydraulic systems.
However, severe pressure fluctuations occur in the switched hydraulic systems, which affect the control
accuracy of the system. Some methods were developed to alleviate the pressure fluctuations for high-precisioncontrol. The proposed digital hydraulic circuit employed the proper duty cycle to control switching valves andrealize an energy-efficient position in [20]. The position control strategy and switching optimization of thedigital valve system is investigated in [21], which not only improves positioning accuracy but also optimizes theswitching number. Although the advanced control theory has improved the dynamic and static characteristics forthe digital switching system, the switching impact is the main issue for switched hydraulic systems and someparameters have a great influence on the output characteristics. Sakai[30] investigated a new structure
preserving nondimensionalization for visualization of hydraulic cylinder dynamics to clarify comprehensive
relations. The sensitivity analysis is recognized as an effective way to quantify the effect of each parameter,
which is beneficial to improve the position accuracy, alleviate pressure fluctuations and optimize control. Due tothe development of modern control theory, sensitivity analysis is widely accepted in many fields to study the1

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variables that the influence of parameters upon system performance [22-24]. According to recent studies[25-27],
many experts research the valve-controlled system by sensitivity analysis to find the main parameters affecting
system characteristics and identify the effect of interactions between variables, such as hydraulic modeling andnonlinearity[31], hydraulic element design.
To achieve the aims outlined above, this paper analyzes the main parameters affecting the system performanceof the MPSCS for the controller design and improvement of output characteristics. The outline of this paper isorganized as follows: In Section II, the principle and mathematical modeling of MPSCS are described forsensitivity function. In Section III, the first-order tracking sensitivity function is solved and each parameter isanalyzed by simulation. In Section IV, the sensitivity analysis theory of MPSCS is validated by experiments.
Finally, conclusions are present in Section V.

SYSTEM DESCRIPTION AND MODELLING
Principles of MPSCSMPSCS [19] mainly consists of accumulator charging circuits, on/off valve matrix, hydraulic transformer,
proportional valves and actuators, etc. (Fig. 1). The valve matrix is a rectangular array, arranged in m and n, (m≥ 2, n≥ 2). The intersections of rows and columns are usually set in on/off valves. There are many connection
methods between the valve matrix and other components (pumps, accumulators and actuators) for different
functions. The pumps are only used to supply the average power demand, while the hydraulic accumulators areused for peak power requirements and recovering the surplus energy. In addition to the tank pressure rail, the
MPSCS can be classified into p1, p2… pn. Each pressure level needs an accumulator to maintain the systempressure. The different pressure rails are introduced to minimize the throttling losses across the valves andenable energy recuperation by switching the valve matrix. In order to balance the energy among the differentpressure levels, the hydraulic transformer is necessary to accomplish the energy transformation
p1

p2

p3

v

On/off valve
matrixpn

Cylinder
FL

6
p1

...

m

p2

q1

q2

SV1

...

3

...

2

...

1

...

...

...

...

...

7.1

...

...

...

...

m-1

DT1

2.1

2.2

DT2

pi
DT3

2.3

SV2
7.2

pj
DT4

2.4

4.1

Hydraulictransformer

Pumps(a) Composition of MPSCS.

DT7
2.7

1.1

DT6

2.5

TP

MP

5.1

...

2.5

4.2

HP

Actuators
Control Unit

DT5

5.2

DT8
2.8

1.2
3.1

3.2

(b) Principle of Three-level Pressure Switching Control System.
Figure 1. Schematic Diagram of MPSCS.

In this study, we take one valve-controlled asymmetric cylinder hydraulic switching system with three-levelpressure as the research object. The pumps and accumulators are installed in the columns, and the actuators areconnected to the rows. The two chambers of the cylinder can connect independently to any oil circuit of thethree-level pressure by on/off valves for different combinations of pressure.

State Equations
To obtain the main parameters influencing the output characteristics of the MPSCS, the mathematical models of
the MPSCS were built in detail [28-29]. Based on the model of state variables of modern control theory, all thevalues of state variables at any time can be obtained, which is helpful to understand the internal state and its
dynamic changes and obtain the dynamic characteristics of the system. We select four state variables that arelinear irrelevance, x1 =y , x2 =y , x3  p1 , x4  p2 . The friction Ff, load force FL, pressure of proportional valves
pi and pj, and the input voltage of proportional valves Ug1 and Ug2 are defined as the input variables u. This

2

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paper investigates the displacement control characteristics of the MPSCS, so the displacement y is defined as theoutput variable. The state equations were deduced in Eq. (1).

 x1  x2
 x   K x  B p x  sgn  x  A1 x  sgn  x  A2 x  Ff  FL
1
2
2
3
2
4
 2
mt
mt
mt
mt
mt


 sgn  x2  A1 x2  Cip  Cep  sgn  x2  K c1  x3 
e


x 


3

Vg1  A1 L0  A1 x11 Cip x4  sgn  x2  K q1 K svU g1  K c1 pisgn  x2  A2 x2  Cip x3  Cip  Cep  sgn  x2  K c 2  x4 
e


x 


 4 Vg 2  A2  L-L0   A2 x1  sgn  x2  K q 2 K svU g 2  K c 2 p j




(1)

It involves many system parameters, some are the structural parameters and others are the working parameters.
Let us introduce them in detail. Yr represents the displacement input signal. The amplification gain Ksv and theflow gain Kq1 or Kq2 of the proportional valves is regarded as a constant value. And the Kc1 and Kc2 are the flowpressure coefficient of the proportional valves. The pi and pj are the pressure of proportional valves for therodless and rod cavity of the cylinder. The p1 and p2 represent the pressure of rodless and rod cavity. The A1 andA2 are the rodless and rod area of the cylinder cavities. The Cip and Cep are the internal and external leakage
coefficient of the cylinder, respectively. The stroke of the hydraulic cylinder piston is L, and the initial positionof the piston is L0. Vg1 and Vg2 are the channel volume from the rodless and rod cavity to the proportional valve.
The βe is the effective volume modulus subjected to gas content and temperature of the oil. The Ky is theconversion coefficient between actual output displacement Y and displacement sensor voltage Up. The mt is thesum mass converted to the rod, and the K is the load stiffness. The Bp is the viscous damping coefficient. The FL
and Ff represent the load force and the friction, respectively.
Therefore, the state equation and output equation are expressed as follow,

 0

 K
 x1    m
x   t 2=
 x3   0
  
 x4  
 0


0
 1

  mt

0
0




1
Bp

0
sgn  x2 mt

 sgn  x2  A1
sgn  x2  A2
0

0
0

1
mt

e
V1

A1
mt

 Cip  Cep   e
 sgn  x2    K c1
 V1

e

Cip

V2

0

0 0

0

0 0

 K c1
Kc2

0 K q1 K sv
0 0



A2

 sgn  x2 
  x1 mt x 
 2
Cip  x3  x 
 4
 Cip  Cep e   sgn  x  K  V 2c2  2 0(2)

 Ff   FL  p 
0
 i   pj 0U  K q 2 K sv   g1 U g 2 0

 x1 x y  1 0 0 0  2 
 x3 
 
 x4 

(3)

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FIRST-ORDER TRAJECTORY SENSITIVITY SOLUTION
Sensitivity Equation
To analyze the influence of parameters on output displacement characteristics, the state equation of MPCSC canbe rewritten as,

 x1   f1 ( x,u,α, t)  x   f ( x,u,α, t) x  f ( x,u,α, t )   2  =  2
  

  

 xn   f n ( x,u,α, t) 

(4)where x  [ x1 ,x2 ,x3 ,x4 ] , u  [u1 ] , α  [α1 ,α2 ,α3 ,α4 ,α5 ,α6 ,α7 ,α8 ,α9 ,α10 ,α11 ,α12 ,α13 ,α14 ,α15 ] , α1  K q1 , α2  K c1 ,α3  Kq 2 , α4  K c 2 , α5  Cip , α6  βe , α7  pi , 8  p j , α9  L , 10  L0 , α11  mt , α12  K , α13  Bp , α14  FL ,α15  K sv .
There are fifteen key parameters that cover the structure parameters, working parameters and control parameters,
take the partial to α for the first-order trajectory sensitivity model.
 f1 ( x,u,α, t) 


 i


 f 2 ( x,u,α, t) x
f ( x , u ,α ,t ) 
x
f
 f
 i
(
)n 

  ( x ) n  (  ) n  (  ) n , ( n  4)
 i
 i
i
i




 f n ( x,u,α, t) 


 i



 f1
 x
 1
 f 2f
where, ( ) n   x1
x f n
 x
 1

f1
x2
f 2
x2
f n
x2

f1 
xn 

f 2  xxxn  , ()n   1
 i  if n 
xn 

x2
 i

T

 f
xn 
f
)n   1
 , ( i  i  i

(5)

f 2
 i

T

f n 
 .
 i 

The sensitivity function of the state vector to the parameter vector is defined as,

λ0i  (

x0
)n
αi

(6)

Then, the sensitivity equation of the MPSCS can be expressed as,λi  (

f
f)n λ i  (
)n
x
i

(7)

where (f / x ) n is the coefficient matrix, and (f /  i ) n is a free item matrix.
Jacobean matrix (f / x ) n can be expressed as follow,

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 a1,1af2,1xn  a3,1
 a4,1

a1,2
a2,2
a3,2
a4,2

a1,3
a2,3
a3,3
a4,3

a1,4 
a2,4 
a3,4 

a4,4 

(8) sgn  x2  A1 x2



 
K
2 
0,  ,  A1 βe (Vg1  A1 L0  A1 x1 )  Cip  Cep  sgn  x2  K c1  x3
 ,
mt



Cip x4  sgn  x2  K q1 K svU g1  K c1 pi  f where, ,x1 
sgn  x2  A2 x2  Cip x3





2 

 A2 βe Vg 2  A2 ( L  L0 )  A2 x1   Cip  Cep  sgn  x2  K c 2  x4 





 sgn  x2  K q 2 K svU g 2  K c 2 p j 


T

T

 Bp

,  sgn  x2  A1 βe (Vg1  A1 L0  A1 x1 ) 1 ,
1, 
f
mt

 ,
x2 
1

sgn  x2  A2 βe Vg 2  A2 ( L  L0 )  A2 x1 

T

A1

1 
0,sgn  x2  m ,  βe Cip  Cep  sgn  x2  K c1  (Vg1  A1 L0  A1 x1 ) ,
f
t

 ,
x3 
1



C
β
V

A
(
L

L
)

A
x
2
0
2 1
 ip e  g 2

T

A2


1
0,  sgn  x2  m ,Cip βe (Vg1  A1 L0  A1 x1 ) ,

f
t

 .
x4 
1 



  βe Cip  Cep  sgn  x2  K c 2  Vg 2  A2 ( L  L0 )  A2 x1  
Jacobean matrix (f /  i ) n can be expressed as follow,

b1,1 b1,2
b
 2,1 b2,2f i bn,1

b4,1b1,ibn,ib1,15 







b6,15 (9)where,TTff  0 , 0 ,sgn  x2  βe K svU g1 (Vg1  A1 L0  A1 x1 ) 1 , 0  ,
 0, 0, βe  sgn  x2  x3  pi  (Vg1  A1 L0  A1 x1 ) 1 , 0 
1
 2
1 T
f
 0 ,0 ,0 ,  sgn  x2  K svU g 2 βe Vg 2  A2 ( L  L0 )  A2 x1   , 3 
1 T
f  0 , 0 ,0 ,  βe  sgn  x2  x4  p j  Vg 2  A2 ( L  L0 )  A2 x1   , 4 

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1 T
f
 0 ,0 , βe   x3  x4  (Vg1  A1 L0  A1 x1 ) 1 , βe  x3  x4  Vg 2  A2 ( L  L0 )  A2 x1   , 5  sgn  x2  A1 x2





1
0 ,0 ,  Cip  Cep  sgn  x2  K c1  x3
 (Vg1  A1 L0  A1 x1 ) ,






Cip x4  sgn  x2  K q1 K svU g1  K c1 pi 
f

 ,
 6  sgn  x2  A2 x2  Cip x3



 

1
  Cip  Cep  sgn  x2  K c 2  x4  Vg 2  A2 ( L  L0 )  A2 x1   sgn  x2  K q 2 K svU g 2  K c 2 p j 

T

T
1 T
f
f
 0, 0,  K c1 βe (Vg1  A1 L0  A1 x1 ) 1 , 0  ,
 0 ,0 ,0 ,  K c 2 βe Vg 2  A2 ( L  L0 )  A2 x1   ,
 7
 8 

sgn  x2  A2 x2  Cip x3


 
2 
f

 0 , 0 , 0 ,  A2 βe Vg 2  A2 ( L  L0 )  A2 x1   Cip  Cep  sgn  x2  K c 2  x4   , 9  sgn  x2  K q 2 K svU g 2  K c 2 p j  T sgn  x2  A1 x2



 

0 ,0 ,  A1 βe (Vg1  A1 L0  A1 x1 ) 2  Cip  Cep  sgn  x2  K c1  x3
 ,




Cip x4  sgn  x2  K q1 K svU g1  K c1 pi  
f

 ,
10 
sgn  x2  A2 x2  Cip x3





2 

 A2 βe Vg 2  A2 ( L  L0 )  A2 x1   Cip  Cep  sgn  x2  K c 2  x4 





 sgn  x2  K q 2 K svU g 2  K c 2 p j 


T

Bp
Ff  FL
 K
 f


A
A
f
1
 0 , 2 x1  2 x2  sgn  x2  12 x3  sgn  x2  22 x4 
, 0, 0  , 0,  x1 , 0, 0  ,2
11  mt
mt
mt
mt
mt
mt

 12 T

TTT
0,0,sgn  x2  K q1U g1 βe (Vg1  A1 L0  A1 x1 ) 1 ,


 f

 f
f
1
1

 0,  x2 , 0, 0  , 0,  , 0, 0  ,1
13 
mt
mt
 14  15   sgn  x2  K q 2U g 2 βe Vg 2  A2 ( L  L0 )  A2 x1  T

It can be considered that at the initial moment, the displacement, velocity and pressure of the hydraulic cylinder
are zero, so the initial value of x is x0=0. Then, the initial value of the sensitivity function  i can be expressed
as follow,

λ0i  0（，i  1...15）(10)

SENSITIVITY ANALYSIS
Sensitivity Indicators
Due to the changing state of the MPSCS in the sampling time, and the value λ changes with time as well. Then,the curve of change is defined as a sensitivity-time history curve, which describes the changing process of
sensitivity function. The change of parameters vector α would cause a change in the state vector x , thus we
could get the Taylor expansion, x 
Δ x    Δ α  o α α n(11)

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It can be seen that the small changes of x will be calculated as long as multiplying the calculated  x /  n by
Δα. There are two indicators[32] to reflect directly the effects of interactions among on state vector x.
For displacement step control of MPSCS, both given displacement and initial displacement are the constantvalue. Therefore, the peak sensitivity S1 is the percentage for the change of the state vector relative to thedifference between the given displacement and the initial displacement, reflecting the impact degree.
S1 

Δ xj
xsj

100% λij Δ αi
xsj

100%

(12)

That is, mean sensitivity S2 represents the overall degree of influence of parameter vector change on the step
response to displacement in the dynamic response process.
t

S2   (ox) n Δ  dt


(13)

This paper selects the step displacement response to analyze the sensitivity, which is easy to observe the rapidity
and stability of the MPSCS. And the peak sensitivity and the mean sensitivity of each parameter are calculated
and compared in the system to reflect the quantitative analysis results of the influence degree of each parameterchange in the system on the displacement output of the system.

Sensitivity Analysis
Based on above sensitivity theoretical model of the MPSCS, the solution of the sensitivity equations need to besynchronously calculated for the real-time variation of the state variables in the system. According to the Eq. (8),there are n×(i+1) expressions to solve, including n expressions of state equations and n×i expressions ofsensitivity equations. We only investigate the sensitivity of x1 for parameters α. Firstly, the equation ofsensitivity function was solved and the programmed Runge-Kutta run in MATLAB/Simulink. Then, thesimulation model of the MPSCS is built based on the mathematical model, and the detailed simulation
parameters are listed in the Table 1.
Table 1. Simulation Parameters of the MPSCS.
Value
100 mm
45 mm
400 mm176 Kg40 Hz
0.707
60 L/min
7.85×10-3 m2
6.26×10-3 m2
40 L

Parameters Name
Diameter of Cylinder D
Diameter of Rod d
Stroke L
Load Mass mt
Frequency of Proportional ValvesDamping ratio of Proportional Valves
Flow of Proportional Valves
Area of Rodless Cavity A1
Area of Rod Cavity A2Accumulator Volume

With the difference of the external load force and step displacement, the system response characteristics of the
MPSCS in actual work are different. The sensitivity of the system parameters is studied under different load
force to analyze the system. The influence of various parameters on the system displacement output
characteristics under different load conditions makes the sensitivity analysis results more applicable. The
rapidity and stability of the system can be judged more conveniently by the step displacement curve, this paper
obtains the sensitivity of each parameter under 5 mm and 10 mm step displacement. The sensitivity-time history
curves λ1i of displacement x1 to αi were obtained under 5 mm step displacement and 10 kN load force in Fig.2.
Similarly, in other cases, it won't go into detail.

7

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8
6
4
2
0
-2
0.8

1.0

-3

0.25
0.00
-0.25
-0.50
-0.75
-1.00
-1.25
0.0

Time t/s

pa)/m ]The sensitivity function of  /×10 [(s·
3

0.0

-2.5

0.6

0.8

1.0

Time t/s(m/pa)
-10

2
1
0
-1
-2
-3

0.2

0.4

0.6

0.8

1.0

(g) Sensitivity Function of λ17

1.5
1.0
0.5
0.0
-0.5
0.0

0.2

0.4

0.8

15
10

5
0
-5
0.4

0.6

0.8

1.0

4.5
4.0
3.5
3.0
2.5
2.0
1.5
1.0
0.5
0.0
-0.5
-1.0
0.0

0.2

0
-4
-8
-12
0.0

0.2

0.4

0.6
Time t/s

0.8

1.0

2

1

-1
-2
-3
-4
0.0

0.2

0.4

0.6

0.8

1.0

0
-5
-10
-15
-20
0.0

0.2

6
4
2
0
-2
-4
0.0

0.2

0.4
0.6
Time t/s

0.8

1.0

0.6

0.8

1.0

2
1
0
-1
-2
0.0

0.2

0.4

0.6

0.8

1.0

(l) Sensitivity Function of λ1

12

2.5
15

The sensitivity function of 1 /V

8

0.4

Time t/s

10

-8

10

1.0

-8

0

15
The sensitivity function of 11410 /(m/N)

2-9The sensitivity function of 11310 /[m /(N·s)

12

0.8

(i) Sensitivity Function of λ19

11

14

0.4
0.6
Time t/s

Time t/s(k) Sensitivity Function of λ1

10

1.0

5

4

Time t/s(j) Sensitivity Function of λ1

0.8(f) Sensitivity Function of λ16

8

-6

20

0.4
0.6
Time t/s

1.0

The sensitivity function of 112/×10 (m /N)

The sensitivity function of 111/×10 (m/kg)-3

The sensitivity function of 110 /×10

0.6

0.2

(c) Sensitivity Function of λ13(h) Sensitivity Function of λ18

25

0.2

-6
0.0

2.0

Time t/s

-10
0.0

-4

(e) Sensitivity Function of λ15

The sensitivity function of 18 /×10

The sensitivity function of 17 /×10-10(m/pa)(d) Sensitivity Function of λ14

-4
0.0

1.0

2.5

5
1

4

2.5

0.4

0.8

0
-2

3.0

2

pa)/m]The sensitivity function of 1 /[(s·

5.0

0.2

0.4
0.6
Time t/s2(b) Sensitivity Function of λ12(a) Sensitivity Function of λ11

-5.0
0.0

0.2

4

(m/pa)

0.6

0.50

-12

0.4

6

The sensitivity function of 16/×10

0.2

0.75

-3

-4
0.0

8

The sensitivity function of 19 /×10

1

10

1.00

The sensitivity function of 13 /×10 (s/m)

12

1.25

pa)/m]The sensitivity function of 12 /[(s·The sensitivity function of 1 /×10-3 (s/m)

14

5
0
-5
-10

2.0

1.5
1.0
0.5
0.0

-15
0.0

0.2

0.4
0.6
Time t/s

0.8

1.0

0.0

0.2

0.4

0.6

0.8

1.0

Time t/s(m) Sensitivity Function of λ1(n) Sensitivity Function of λ1(o) Sensitivity Function of λ1
Figure 2. The Sensitivity-Time History Curves of Parameters
14

13

15

It can be seen from Fig.2 that each parameter influences output displacement x1. The positive and negative mean
growth and inhibition, respectively. When the displacement x1 approaches the steady-state, the sensitivity
functions of the parameters α7, α8, α13, α14 have a non-zero solution and is not close to zero, which have great
influence on the state characteristics of displacement. To visually observe the influence of these parameters,
dynamic displacement characteristics of 5 mm and 10 mm step under 10 kN and 20 kN loading force wereanalyzed when 10% parameters change by sensitivity indicators. The pre-pressure of rodless or rod cavity is the8

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The 11th JFPS International Symposium on Fluid Power 2020

switching pressure level, the parameter change in pressure is the change of pressure level actually. The
parameters α7, α8 would change simultaneously, thus the two parameters merge into one. The results ofsensitivity analysis is shown in Fig.3 and Fig.4.
0.3

0.0

0.1

0.0

1 2 3 4 5 6 7,8 9 10 11 12 13 14 15
Parameters number

4

0.4

3

0.3

2

1

0
1

2

3

4

5

6 7,8 9 10 11 12 13 14 15

Parameters number(a) Peak Sensitivity with
5 mm Step Displacement

1.0

0.5

0.0

1 2 3

1

4 5 6 7,8 9 10 11 12 13 14 15
Parameters number

2

3

4

5

0.1

0.0

0.05

0.00

6 7,8 9 10 11 12 13 14 15

1

1.0

0.5

0.0
1

2

3

4

5

6 7,8 9 10 11 12 13 14 15

Parameters number

Parameters number

3

4

5

6 7,8 9 10 11 12 13 14 15(d) Mean Sensitivity with
10 mm Step Displacement

0.20

1.5

1 2 3 4 5 6 7,8 9 10 11 12 13 14 15

2

Parameters number

2.0

0.2

0.10

Parameters number(b) Mean Sensitivity with(c) Peak Sensitivity with
5 mm Step Displacement
10 mm Step Displacement
Figure 3. Sensitivity Indicators under 10 kN Load Force

Mean sensitivity S2/mm

Peak sensitivity S1(%)(a) Peak Sensitivity with
5 mm Step Displacement

Mean sensitivity S2/mm

0.5

0.2

0.15

Mean sensitivity S2/mm

1.0

Peak sensitivity S1(%)

1.5

1.5

Peak sensitivity S1(%)

2.0

Mean sensitivity S2/mm

Peak sensitivity S1(%)

2.5(b) Mean Sensitivity with(c) Peak Sensitivity with
5 mm Step Displacement
10 mm Step Displacement
Figure 4. Sensitivity Indicators under 20 kN Load Force

0.15

0.10

0.05

0.00

1 2 3 4 5 6 7,8 9 10 11 12 13 14 15
Parameters number(d) Mean Sensitivity with
10 mm Step Displacement

According to the results of sensitivity analysis, the peak and mean sensitivity indicators of the parameters 1, 2, 3,
4, 5, 6, 9, 10, 11, 12 and 13 are very smaller value compared with the parameters of 7, 8, 14 and 15. The fourparameters have a significant influence on the step displacement response in the dynamic characteristics. Under
the 10 kN load force, the interaction of parameters 7 and 8 have the most influence on the displacement. The
values of peak sensitivity of 5 mm and 10 mm displacement step are 2.24% and 1.15%, respectively. The valuesof mean sensitivity are 0.28 mm and 0.15 mm, respectively. However, the values of peak sensitivity of 5 mm
and 10 mm displacement step with 20 kN are 3.57% and 1.73%, respectively. While, the values of meansensitivity are 0.30 mm and 0.15 mm, respectively. Therefore, the values of peak sensitivity have a certain
increasing trend with the load force increasing under the same displacement step, while the values of meansensitivity have little change. Besides, both peak sensitivity and mean sensitivity decreased to some extent withthe same load force. The parameter 14 is load force, the trend of the change of two sensitivity indicators are thesame. The values of the two sensitivity increase dramatically under the same displacement step, while the values
decrease under the same load force. For parameter 15, under the different displacement steps, the values of peaksensitivity under 10 kN load force are 0.48% and 0.50%, the values of mean sensitivity are 0.11 mm and 0.10
mm. Under 20 kN load force, the value of S1 are 0.68% and 0.65%, the value of S2 are 0.19 mm and 0.16 mm.
Hence, the values of two indicators have a little change with the change of displacement step, while it has an
obvious increase with load force rising. In conclusion, pre-pressure parameters, load force and proportional
action factor have greater influence parameter on the displacement output, especially the pre-pressure.

EXPERIMENT VERIFICATIONTo ensure the accuracy and reliability of the sensitivity analysis results, it is necessary to carry out experimental
verification on the analysis results. However, many parameters of the MSPCS are difficult to measure, andachieve the experimental verification of the sensitivity analysis results of all parameters. Therefore, this paper
verifies the sensitivity analysis results by some measurable parameters through experiments, and prove theanalysis conclusions of other parameters by analogy. The parameter 7 and 8, which are changed by switching
the pressure levels, the parameter 10 and the parameter 14 are selected to verify the sensitivity theory analysis
by experiments.

9

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Experiment Platform
The experiments were carried out on the MSPCS platform, which mainly includes the hydraulic transmission
system and the data acquisition and control system, see Figure 5. The servo valve-controlled cylinder system offorce closed-loop control is adopted to simulate load force. The multi-level pressure achieve by the connection
between the actuator and the pressure level by opening and closing of the corresponding on/off valve in thecontrol valve matrix. Each pressure level is composed of an accumulator with an unloading relief valve and amanual variable pump, the maximum safe working pressure set by the unloading relief valve to compare withpressure level to decide the operation or unloading of the manual variable pump.
The signal control and acquisition system for MPSCS is designed on MATLAB/Simulink software in a personal
computer (PC), and the control signals are sent to an industrial personal computer (IPC) to control actuator
directly in real-time. Control signals from IPC are sent to the on/off valve matrix to select the corresponding
pressure levels. At the same time, the IPC is responsible for signal acquisition, thus the collected pressure anddisplacement data could processed by the real-time controller with 1 ms sample interval. Moreover, the userinterface could help to monitor the data and control process, as well as adjust some system parameters. The
components and used transducers are listed in Table 2.
Accumulators

xPC
On/off valves
matrixSignalCondition

Actuator

Pumps
Simulator of load

Host PC
Servo
AmplifierForce sensor
Displacement sensor

UPS Power(a) Hydraulic Transmission System(b) The Signal Control and Acquisition System
Figure 5. MPSCS Platform
Table 2. Main Parameters of Experimental Components
Parameters Name
Value

Manual variable pump
Proportional variable pump(MOOG-RKP)
Accumulator
CylinderHydraulic motor
Displacement sensor
Force sensor
Speed torque sensor
Pressure sensor
Flow sensor

0-31.5 MPa；0-25 mL/rev
0-28 MPa；19 mL/rev
40 L∅100 mm/45 mm-400 mm
940 mL/rev
400 mm；0.1%FS
20 kN；0.05%FS
2000 r/min；5000 Nm；0.5%FS
25 MPa；0.5%FS
60 L/min；2%FS

Results and Discussion
The H pressure level is set as 60 bar, and the M pressure level is set as 30 bar. The initial position of rod L0=200
mm. Besides, the 5 mm and 10 mm displacement step under 10 kN and 20 kN load force were given and shownin Fig. 6 and Fig.7.

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Reference

212

Experiment

205
204

Displacement x1 /mm

Displacement x1 /mm

206

203
202
201
200
199
198
0.0

0.2

0.4
0.6
Time t/s

0.8

Experiment

208

206
204
202
200
198
0.0

1.0

Reference

210

0.2

0.4
0.6
Time t/s

0.81.0(a) Step Response of 5 mm under 10 kN(b) Step Response of 10 mm under 10 kN
Figure 6. Displacement Step Response under 10 kN
Reference

212

Experiment
Displacement x1 /mm

Displacement x1 /mm

206
205
204
203
202
201
200
199
198
0.0

0.2

0.4
0.6
Time t/s

0.8

Experiment

208
206
204
202
200
198
0.0

1.0

Reference

210

0.2

0.4
0.6
Time t/s

0.81.0(a) Step Response of 5 mm under 20 kN(b) Step Response of 10 mm under 20 kN
Figure 7. Displacement Step Response under 20 kN

With 10 % variations of above parameters, the comparison between the experiment and theoretical simulation
analysis of two indicators are shown in Fig.8
5 mm
10 mm

2
1

10 kN
Experiment

20 kN
Experiment

0.3

0.2

10 kN
Experiment

20 kN
Theory

0.2

0.1

10 kN
Theory20 kN
Experiment

20 kN
Theory20 kN
Experiment

5

2
1

0.6
0.4
0.2

10 kN
Experiment

20 kN
Experiment

20 kN
Theory20 kN
Experiment

0.8

20 kN
Theory

5 mm
10 mm

0.6

0.4

0.2

0.0

10 kN
10 kN
Experiment Theory

10 kN
Theory(c) Peak Sensitivity of Parameter 10

5 mm
10 mm

3

5 mm
10 mm

0.8

0.0

20 kN
Theory

4

0

0.0

10 kN
Theory(b) Mean Sensitivity of Parameters 7, 8

5 mm
10 mm

Peak sensitivity S1(%)

Mean sensitivity S2/mm

0.4

0.0

10 kN
Theory(a) Peak Sensitivity of Parameters 7, 8

10 kN
Experiment

0.6

Mean sensitivity S2/mm

0Peak sensitivity S1(%)

4
3

1.0

5 mm
10 mm

0.8

Mean sensitivity S2/mm

Peak sensitivity S1(%)

5

10 kN
Experiment

10 kN
Theory20 kN
Experiment

20 kN
Theory(d) Mean Sensitivity of Parameter 10(e) Peak Sensitivity of Parameter 14(f) Mean Sensitivity of Parameter 14
Figure 8. Sensitivity Indicators with 10% Changes of Parameters

The experimental results demonstrate that the two sensitivity parameters almost consistent with the theoreticalanalysis. The parameter 10 has very little influence on the displacement, but it is susceptible to other factors
easily. Thus, the results show the some difference value between simulation and experiment measure. Similarly,

11

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The 11th JFPS International Symposium on Fluid Power 2020the load force also has a difference due to the impact of transient force, which enhances the change range of load
force, thus the experiment indicators of load force is greater than the simulation results. The main reason forvalue deviation is that it is difficult to describe all the state characteristics based on the current mathematicalmodel, which affects the sensitivity analysis of parameters. It makes the unavoidable deviation between theory
and experiment.
Based on above theoretical calculation and experimental verification, the influence of each variation parameteron the output of the MPSCS is quantitatively analyzed. The results show that the front pressure of proportional
valve, load force and amplification coefficient of proportional valve significantly affected the displacement
output characteristics, especially the front pressure of the proportional valve that has the most influence on
output characteristics. This can provide a theoretical basis for MPSCS control optimization and compensation.

CONCLUSION
To address the switching impact and optimize control to obtain better compensation for the Multi-level Pressure
Switching Control System (MPSCS), this paper develops the sensitivity analysis method to quantify the
influence of the main parameters on the MPSCS characteristics. The first-order trajectory models of sensitivity
were established based on the mathematical models of MPSCS and the state equations. We solved the sensitivity
function of the fifteen parameters and got the sensitivity-time history curves of each parameter under 5 mm step
displacement and 10 kN load force based on the simulation model built by MATLAB/Simulink. The parameter
7, 8, 12 and 14 have great influence on the state characteristics of output displacement.
The influence degree of each parameter on the displacement output dynamic characteristics of the MPSCS wasquantitatively analyzed by the proposed peak sensitivity and mean sensitivity. The parameters 7, 8, 14 and 15
have a significant influence on the step displacement response in the dynamic characteristics under 10%
parameters change, especially the parameters 7 and 8. The experiments were carried out on the MPSCS platform
to validate the feasibility of sensitivity theory by some measurable parameters 7, 8, 14. The experimental results
demonstrate that the two sensitivity indicators almost consistent with the theoretical analysis. Though there is asmall deviation between theory and experiment, it is unavoidable due to the current mathematical models are notaccurate entirely to describe all the state characteristics of the MPSCS. Next, the nondimensionalization would
be tried to apply on sensitivity analysis for more general result. Future work will be carried out on the control
optimization and compensation for the MPSCS based on the results of the sensitivity analysis.

ACKNOWLEDGMENTS
The authors thank the anonymous reviewers for their critical and constructive review of the manuscript. This
study was co-supported by the National Natural Science Foundation of China (No.51975507).

REFERENCES
1. PAN, M., PLUMMWER, A., Digital Switched Hydraulics. Frontiers of Mechanical Engineering, 2018, 13-2, pp.225-231.
2. SCHEIDL, R., LINJAMA, M., SCHMIDT, S., Is the Future of Fluid Power Digital Proceedings of the Institution of
Mechanical Engineers. Part I, Journal of Systems and Control Engineering, 2012, 226-6, pp.721– 723
3. YANG, H. Y., PAN, M., Engineering Research in Fluid Power: A Review. Journal of Zhejiang University. Science A,
2015, 16-6, pp.427–442.
4. LINJAMA, M., LAAMANEN, A., VILENIUS, M., Is It Time for Digital Hydraulics? In: Proceedings of the 8th
Scandinavian International Conference on Fluid Power. Tampere, 2003, pp.347–366.
5. LINJAMA, M., Digital ﬂuid power: State of the art. In: Proceedings of the 12th Scandinavian International Conferenceon Fluid Power. Tampere, 2011, pp.18–20.
6. JOHNSTON, D. N., A Switched Inertance Device for Efficient Control of Pressure and Flow, In Proceedings of the
ASME 2009 Dynamic Systems and Control Conference - DSCC2009, Hollywood, USA, 2009, pp.589-596.
7. DE, NEGRI, V. J., WANG, P., PLUMMER, A., JOHNSTON, D. N., Behavioural Prediction of Hydraulic Step-up
Switching Converters, International Journal of Fluid Power, 2014, 15-1, pp. 19.
8. KOGLER, H., SCHEIDL, R., Two Basic Concepts of Hydraulic Switching Converters. In Proceedings of the First
Workshop on Digital Fluid Power, Tampere, Finland, 2008, pp. 113128.
9. PAN, M., HILLIS, A., JOHNSTON, N., Active Control of Fluid-bome Noise in Hydraulic Systems Using In-series andBy-pass Structures, 2014 UKACC International Conference on Control. IEEE, 2014, pp.355-360.
10. VUKOVIC, M., MURRENHOFF, H., The Next Generation of Fluid Power Systems, Dynamics and Vibroacoustics of
Machines, 2014, pp.1-7.

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11. MURRENHOFF, H., SGRO, S., VUKOVIC, M., An Overview of Energy Saving Architectures for Mobile Applications,
In Proceedings of the 9th International Fluid Power Conference, Aachen, Germany, 2014b, pp.374-385.
12. VUKOVIC, M., SGRO, S., and MURRENHOFF, H., STEAM – A Mobile Hydraulic System with Engine Integration, In
Proceedings of the ASME/BATH 2013 Symposium on Fluid Power & Motion Control, FPMC 2013, Sarasota, Florida,USA.
13. HANSEN, R. H., ANDERSEN, T. O., PERDERSEN, H. C., Analysis of Discrete Pressure Level Systems for Wave
Energy Converters, International Conference on Fluid Power and Mechatronics, 2011, pp.552-558.
14. HANSEN, A. H., PEDERSEN, H. C., Optimal Configuration of a Discrete Fluid Power Force System Utilized in the
PTO for WECs. Ocean Engineering, Aalborg East, Denmark, 117, 2016, pp.88-98.
15. HANSEN, A. H., ASMUSSEN, M. F., BECH, M. M., Energy Optimal Tracking Control with Discrete Fluid Power
Systems Using Model Predictive Control. In Proceedings of the Ninth Workshop on Digital Fluid Power, Aalborg,
Denmark, 2017, pp.7-8.
16. LIN, Y., XU, J., LIU, H., et al., Pressure Matching of Wave Energy Device Based on Digital Hydraulic Cylinder Group,
Journal of Zhejiang University (Engineering Science), 2019, 53-10, pp.1892-1897.
17. LEIFELD, R., VUKOVIC, M., and MURRENHOFF, H., STEAM – the Best of Both Worlds, The Seventh Workshop on
Digital Fluid Power, Linz, Austria, 2015.
18. YAO, J., Li, B., KONG, X.D., Energy Saving and Control of Hydraulic Press Fast Forging System Based on the
Two_Stage Pressure Source, Journal of Mechanical Engineering, Vol. 52, No. 10, 2016, p.199-206.
19. YAO, J., CAO, X., WANG,P., et al., Multi-level Pressure Switching Control and Energy Saving for Displacement Servo
Control System, in BATH/ASME 2016 Symposium on Fluid Power and Motion Control, Bath, UK, 2016, pp.7-9.
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Hydraulic System Using on/off Valve, The 4th International Conference on Robotics and Mechatronics (ICROM). IEEE,
2016, pp.234-239.
21. GAO, Q., LINJAMA, M., PALONIITTY, M., et al., Investigation on Positioning Control Strategy and Switching
Optimization of an Equal Coded Digital Valve System, Proceedings of the Institution of Mechanical Engineers, Part I:
Journal of Systems and Control Engineering, 2019.
22. SONG, X. M., KONG, F. Z., ZHAN, C. S., et al., Parameter Identification and Global Sensitivity Analysis of Xin’anjiang Model Using Meta-modeling Approach. Water Science and Engineering, 2013, 6-1, pp.1-17.
23. HALL, J. W., BOYCE, S. A., WANG, Y. L., et al., Sensitivity Analysis for Hydraulic Models, Journal of HydraulicEngineering, 2009, 135-11, pp.959-969.
24. WEI, P.，SONG, J.，LU, Z., Global Reliability Sensitivity Analysis of Motion Mechanisms, Journal of Risk andReliability, 2016, 230, pp.265-277.
25. BA, K., YU B., KONG X., et al., The Dynamic Compliance and Its Compensation Control Research of the Highly
Integrated Valve-controlled Cylinder Position Control System, International Journal of Control, Automation & Systems
Journal of Dynamic Systems, 2017, 15-4, pp.1814-1825.
26. BA, K., YU, B., Parameters Sensitivity Analysis of Position-based Impedance Control for Bionic Legged Robots’ HDU,
Applied Sciences, 2017, 7-10, pp.1035.
27. ANTON, I, LOAN, S., Numerical Modeling of Cavitation Characteristics and Sensitivity Curves for Reversible
Hydraulic Machinery. Engineering Analysis with Boundary Elements, 2014, 41, pp.18-27.
28. CAO, X., GUO, B., WANG, P., et al., Multi-level Pressure Switching Displacement Servo Control Systems, China
Mechanical Engineering, 2017, 28-20, pp.2447-2454.
29. WANG, P., Control Strategy and Design of Multi-Level Pressure Independent Meter-In and Meter-Out Control System,
Yanshan University, China, 2018.
30. SAKAI, S., STRAMIGIOLI, S., Visualization of Hydraulic Cylinder Dynamics by a Structure Preserving
Nondimensionalization, IEEE/ASME Transactions on Mechatronics, 2018, 23-5, pp. 2196-2206.
31. KHALIL, H. K., Nonlinear Systems. Englewood Cliffs, NJ, USA: Prentice-Hall, 1996.
32. NURMI, J. and MATTILA, J., Detection and Isolation of Leakage and Valve Faults in Hydraulic Systems in Varying
Loading Condition, Part 1: Global Sensitivity Analysis, International Journal of Fluid Power, 2011, 12, pp. 41–51.

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A Study on the Pulse Analysis and Vibration Characteristics of Hydraulic
System for Prediction of Check Valve Behavior
Jeong-Woo PARK*, ** Eun-A JEONG*, Young-Bog HAM* and So-Nam YUN**

Department of Thermal Systems, Energy Systems Research Division,
Korea Institute of Machinery & Materials, Daejeon, 34103, Korea.
**
Department of Convergence System Engineering, Chung Nam National University, Korea.
(E-mail: ysn688@kimm.re.kr)

Abstract. The check valve is a basic type of valve. The poppet shaped check valve has been studied for decades andis still a topic of discussion to improve performance. However, most of the papers were based on experiments topredict valve performance and theoretical parts such as valves or pipelines, but few papers evaluated hydraulicvibration characteristics by simultaneously analyzing pressure-flow pulsation from pumps. Therefore, in this study
we propose an experimental circuit that simulates the action of the check valve by generating positive pressure
through a well-controlled pump system.
Keywords: Check Valve, Flow Pulsation, Pressure Fluctuation, Set Pressure.

INTRODUCTION
In order to predict the behavior of the check valve, a model of hydraulic circuit was devised to investigate thepulsation and vibration characteristics of the hydraulic system. The speed of the motor connected to the circuit is
set to 100 [rpm], and the overflow valve was set to 10 [bar] for consecutive set pressure values, so that pressurechanges can be interpreted. The pulsation of pressure and flow was obtained from the inlet of the hose and theinlet of the check valve. The behavior of the check valve was analyzed from the data obtained.

VALVE SYSTEM ANALYSIS AND RESULT GRAPHS
Fig. 1 shows a circuit used to perform experiment for finding characteristics of the check valve. The check valve(No. ①) was installed and the experimental hydraulic unit (No. ②) was used to supply hydraulic oil to thecheck valve. In this study, hydraulic pump (No. ③) installed in Fig. 1 (No. ②) was modeled and analyzed to
predict valve behavior by generating a load at the check valve. Based on Fig. 1 circuit, the Simulation X tool
was used to complete the modeling was divided into pump, hose and valve parts. The results were used tocompare the pressure pulsation and flow pulsation at the front end of the hose with the pressure and flow
pulsation in the check valve.

Figure 1. Hydraulic circuit for behavior prediction of check valve

Figures numbered 2 to 5, shows the pulsation change when the pressure of the overflow valve is set to 10 [bar],

1

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The 11th JFPS International Symposium on Fluid Power 2020

20 [bar], 30 [bar], 40 [bar] and 50 [bar]. Fig. 2 and Fig. 4 show the difference between the minimum andmaximum values of normal pressure pulsation and Fig. 3 and Fig. 5 show the difference between the minimum
and maximum values of normal flow pulsation.
Fig. 2 shows a pressure difference of 0.32 [bar], 2.56 [bar], 5.91 [bar], 9.82 [bar] and 13.48 [bar] for consecutive
set pressure values of 10bars to 50 bar. It can be predicted that the greater would be the set pressure, the greater
the difference in pressure pulsation.

Figure 2. Pressure fluctuations amplitude versus set pressure at the hose inlet

Fig. 3 is a comparison of the magnitude of the flow pulsation difference for different pressure values. It can beseen that there are differences of 0.66 [lpm], 4.33 [lpm], 6.22 [lpm], 7.60 [lpm] and 8.96 [lpm] for consecutive
set pressure values. Thus, the greater the pressure, the greater the pulsation of the flow rate. It can be seen that as
set pressure is increased by 10 [bar], the average increase in flow rate difference turned out to be about 2[lpm].

Figure 3. Flow pulsation amplitude on set pressure across the hoseFig. 4 shows the pressure pulsation difference in the orifice. It can be seen that 0.52 [bar], 3.61 [bar], 6.66 [bar],
10.14 [bar] and 13.81 [bar] remain for consecutive set pressure values. Therefore, it can be said that thevibration from the valve appears as vibration in the pressure pulsation.

Figure 4. Pressure fluctuations amplitude versus set pressure at the check valve

Fig. 5 shows the difference in flow pulsation in the orifice, and summarizes the difference in flow pulsation by

2

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The 11th JFPS International Symposium on Fluid Power 2020pressure. Based on the difference in flow rate of 0.09 [lpm], 0.38 [lpm], 1.06 [lpm], 1.55 [lpm] and 1.71 [lpm]for consecutive set pressure values, which means the higher the pressure, the greater the flow pulsation. Also
because of the hose the pulsation characteristics are better than Fig. 3.

Figure 5. Flow pulsation amplitude on set pressure across the check valve

CONCLUTION
The purpose of this work was to investigate the effect of pressure-flow pulsation on the pump in hydraulicsystem to predict check valve behavior. The results are summarized as follows.
1. As a method proposed in this study, it is confirmed that the pressure-flow pulsation of the hydraulicpump can be copied and the pressure-flow pulsation increases with the circuit setting pressure.
2. It is considered necessary to use an appropriate hydraulic hose in order to clearly replicate the action of
the pre-fill valve in the future because the pulsation generated by the pump varies depending on thepipeline conditions and whether or not the hydraulic hose is installed.

ACKNOWLEDGMENTS
This paper was supported by the Korea Institute of Industrial Technology (P0013851, 2021 Commercial Vehicle
Industry Innovation Growth and Future Industrial Ecosystem Construction Project).

REFERENCES
1. J. E. FUNK, Poppet Valve Stability, Journal of basic Engineering ASME, pp.207-212, 1964.
2. A. AMINI, I OWEN, A Practical Solution to the Problem of Noise and Vibration in a Pressure-Reducing Valve, Experimental
Thermal and Fluid Science, 0894-1777, pp. 136-141, 1995.
3. C. LI, F. LIEN, E. YEE, M. DONG, Multiphase Flow Simulations of Poppet Valve Noise and Vibration,
SAE paper 2015-01-0666, 2015.

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The 11th JFPS International Symposium on Fluid Power 2020

Research on an Oil-hydraulic Component
to Reduce Pressure Pulsation
Yasuo SAKURAI*, Misaki HASHIMOTO**, Moritaka MAEHARA***, Norikazu HYODO***** Department of Innovative Engineering
Ashikaga University268-1 Oomaecho, Ashikaga, Tochigi, 326-8558 Japan(E-mail: ysakurai@ashitech.ac.jp)** MITSUBA Corporation
1-2681 Hirosawa-cho, Kiryu-shi, Gunma 376-8555, Japan*** SAWAFUJI ELECTRIC CO., LTD.
3, Nittahayakawa-cho, Ota City, Gunma 370-0344 Japan**** TOKYO KEIKI INC.
Sakaecho, Sano-shi, Tochigi, 327-0816 Japan

Abstract. Pressure pulsation in oil-hydraulic system is generated due to the use of a positive displacement pump andit causes oscillation, noise and so on. An accumulator is generally employed to suppress the pressure pulsation.
However, accumulator needs regular maintenance to prevent its performance deterioration due to gas leakage from it.
Furthermore, there is a case where it is difficult to make the size of oil-hydraulic system compact due to the size ofaccumulator. Therefore, it seems to be necessary for innovation of an oil-hydraulic system to develop a new oilhydraulic component to reduce pressure pulsation, and then its size is compact and regular maintenance to keep its
performance is not necessary. In this study, an oil-hydraulic component to reduce pressure pulsation is proposed.
The component is fabricated and some experiments are carried out to verify the mechanism to reduce pressure
pulsation and to make clear its basic characteristics.
Keywords: Oil-hydraulics, Component, Pressure pulsation, Rubber

INTRODUCTION
In oil-hydraulic system with positive displacement pump, the pressure pulsation is generated [1], [2] and it
causes vibration and noise [3]-[5]. For example, it is necessary to suppress pressure pulsation in order to finish
surface of metal product precisely in a machine tool [6]. Hence, it is important to reduce the pressure pulsation
so as to make the quality of an oil-hydraulic system higher.
To make the pressure pulsation smaller, an accumulator filled with pressurized gas is usually employed.
However, it is necessary to carry out regular maintenance of the accumulator to prevent its performance
deterioration caused by gas leakage [7]. On the other hand, it seems to be significant to reduce the size ofaccumulator when making the size of oil-hydraulic system more compact. To cope with these problems, it isdesirable to develop a new oil-hydraulic component to reduce pressure pulsation with small size and withoutmaintenance.
To realize this, some of the authors had proposed innovative components using the elasticity of rubber [8], [9].
However, in these components, it is difficult to improve durability, to adjust the gap width which is importantparameter in the performance and to cope with the leakage from the component.
In this study, the component to reduce pressure pulsation is improved, which is able to overcome the abovementioned problems. Some experiments are carried out to confirm the mechanism to suppress pressure pulsationand to investigate the basic characteristics of the improved component. In addition to it, through someexperiments varying the gap width and the hardness of urethan rubber tube, the influence of these design
parameters on the performance of the component are made clear.

MECHANISM TO REDUCE PRESSURE PULSATION
The mechanism of the component to reduce pressure pulsation is shown in Fig.1. The feature of this component
is that the elastic tube is inserted and a part of the tube is exposed to the atmosphere through the gap. When thepressure is increased, a part of the tube expands and prevents the increase of the pressure by the increase of theinside volume of the tube. On the other hand, when the pressure is decreased, a part of it contracts and prevents
the decrease of the pressure by the decrease of its inside volume. Namely, such movement are assumed tosuppress the pressure pulsation.

1

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Vibration

Atmosphere

Atmosphere

Elastic tube
in

Pipe fitting

out

Metal pipe
Gap(a)Basic structure(b)Mechanism
Figure 1. Mechanism of the component to reduce pressure pulsation

STRUCTURE OF IMPROVED COMPONENTThe structure of the improved component is shown in Fig.2. In this study, a urethan rubber tube is employed as
an elastic tube since urethan rubber has excellent resistance to chemicals and has stable characteristics over awide range of temperatures. The inner diameter and thickness of the urethan rubber tube were 24mm and10.5mm, respectively. This component mainly consists of two metal pipes, three pipe spacers, two flanges, aurethane rubber tube and two rubber packings. After attaching the flanges and metal pipes to the pipe fittings,
the urethane rubber tube is inserted into the metal pipes. By attaching the pipe spacer between the flanges, thewidth between the flanges is determined, that is, the length of this component. Owing to adopting this structure,
it becomes easy to determine the position of the metal pipes. By adjusting the position of the two metal pipes,
the gap width between them can be determined between 0mm and 9 mm. The periphery of the metal pipe ischamfered with R8.0 to ensure the durability of the urethan rubber tube by forestalling stress concentration on
the tube when a part of the tube expands. To stop leakage between the urethane rubber tube and the metal pipe,
the rubber packings are newly installed. Its thickness was 2.0mm.
200
170
Pipe spacer

M8
120

Pipe
fitting
Flange

0～9

10.5

φ24

φ 45

Gap

Atmosphere Rubber packing

Metal pipe Urethane rubber tube(a)Assembly drawing
A

17
10

8.2×3

64.5

20

φ75
φ55

Rc2

A

φ45

φ54

Rc1-¼

A-A

(b)Flange(c)Metal pipe(d)Photo of component(e)Photo of parts

Figure 2. Oil-hydraulic component to reduce pressure pulsation2

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PERFORMANCE OF COMPONENT TO REDUCE PRESSUR PULSATION
Confirmation of Mechanism to Reduce Pressure Pulsation
To confirm the mechanism to reduce pressure pulsation, the experiment was carried out when the pumpdischarge pressure was low, that is, 1.5MPa.
The experimental apparatus is illustrated in Fig.3. The experimental apparatus is composed of a piton pump, arelief valve, the component to reduce pressure pulsation, a metering valve and a bubble eliminator to reduce theinfluence of air in oil. The displacement of the pump was 4.0×10-5m3/rev and its angular velocity was 157rad/s(1500rpm). By adjusting the setting pressure of the relief valve and the opening area of the metering valve, theaverage pressure at the pump discharge port set about 1.5MPa and the flow rate through the component set
about 2.5×10-4m3/s (15L/min). Then, oil temperature was set at about 25℃, and both upstream pressure P1 anddownstream pressures P2 of the component were measured by semiconductor strain gauge pressure transducers.
The output voltage from these transducers were recorded on a data logger for 10 seconds with sampling time of
0.1ms.
Pressure
gauge

Piston
pump

Relif
valve

Proposedcomponent
P1
P2

Metering
valve

Pressuretransducer

Flow
meter

Data logger

BubbleEliminator
Tank

Tank

Figure 3. Experimental apparatus

Experimental results are shown in Fig.4. In this experiment, the gap between the two metal pipes was 2mm andthe hardness of the urethane rubber tube was A50. As seen from this figure, the variation width of the pressure
pulsation of the upstream side of the component is about ±0.146MPa, and that of the downstream side of it isabout ±0.071MPa. Namely, the pressure pulsation of the downstream side of the component becomes about
48% of that of the upstream side of it. As a result, it is known that this component is effective to reduce pressure
pulsation. The average pressure at the upstream side of the component and that at the downstream side of it are
1.45 MPa and 1.40 MPa, respectively, and the difference is 0.05 MPa. Therefore, in this experimental condition,
the pressure drop of this component is sufficiently small.
1.6

P2(MPa)P1(MPa)

1.6

1.5

1.5

1.4

1.4
1.3
0

20

40

60

80

1.3
0

100

20

40

60

(a)Upstream(0~100ms)

100(b)Downstream(0~100ms)
1.6

P2(MPa)P1(MPa)

1.6

1.5

1.5

1.4

1.4
1.3
0

80

Time(ms)

Time(ms)

5

10

15

20

1.3
0

5

10

15

20

Time(ms)

Time(ms)(c)Upstream(0~20ms)(d)Downstream(0~20ms)
Figure 4. Experimental results

The results of frequency analysis for the experimental results in Fig.4 are shown in Fig.5. In this experiment, apiston pump was used and its rotational frequency was 25Hz and its piston number was 9. As can be seen from
Fig.5 (a), there are peaks at almost the integral multiplication of frequency 225Hz, which is the product of its
rotational frequency and its piston number. The relation between the frequency and the amplitude are shown in
Table 1. From this table, it is known that the amplitude at 447Hz is 0.55MPa in the upstream side, which is the3

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The 11th JFPS International Symposium on Fluid Power 2020

0.6
0.4
0.2

0
0

0.6

Amplitude (MPa)Amplitude (MPa)largest and the amplitude at 447Hz becomes about 94% smaller in the downstream side. In addition to it, it is
known that the amplitude at other frequencies is also small in the downstream side.
Photos of the urethan rubber tube are shown in Fig.6. As seen from this figure, the urethan rubber tube is
expanded when P1=1.5MPa. Then, displacement X of this part of the urethan rubber tube was measured byusing a laser sensor. The results are illustrated in Fig.7. In this figure, X=0.0mm means the position when thepressure in the tube is 0.0Pa. As can be seen from this figure, it is vibrating around 3.346mm. Hence, it isdeduced that the inside volume of the tube is change due to this vibration and the variation width of the pressure
pulsation of the downstream side of the component becomes small.
Consequently, it becomes clear that the proposed component and its mechanism are effective to reduce thepressure pulsation.

200

400

600

800

Frequency (Hz)

0.4
0.2

0
0

1000

200

(a)Upstream side

400

600

800

Frequency (Hz)1000(b)Downstream side
Figure 5. Frequency analysis resultsFrequency (Hz)
223
447
670
893

Table 1. Frequency and amplitude
Amplitude (upstream) (MPa) Amplitude (downstream) (MPa)
0.11
0.06
0.55
0.035
0.10
6.3×10-4
0.10
4.3×10-4

Metal pipe
Urethan rubber tube

Metal pipe
Urethan rubber tube(a)P1=0.0MPa(b)P1=1.5MPa

Figure 6. Photos of urethan rubber tube
X(mm)

3.350
3.348
3.346
3.344
3.342
3.340
0

40

80

120

160

200

Time(ms)

Figure 7. Displacement of the urethan rubber tube

Investigation of Performance of the Component to Reduce Pressure Pulsation at Practical Pressure
In a machine tool such as an NC lathe, an oil-hydraulic system is often used at 3.5MPa [10]. In this section, theexperimental results are described when the average pressure at the pump discharge port is a practical value, thatis 3.5MPa. On the other hand, it is necessary to investigate the relation between its design parameters andperformance. Therefore, some experiments were carried out by varying the gap width between the metal pipes
and the hardness of the urethan rubber tube.
Experiment was conducted by using the experimental apparatus shown in Fig.3. Then, the average pressure atthe pump discharge port was set about 3.5MPa by adjusting the relief valve setting pressure and the flow rate

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through the component was set about 2.5×10-4m3/s (15L/min) by adjusting the opening area of the metering
valve.
The typical experimental results are illustrated in Fig.8, where the gap between the two metal pipes is 4.0mm
and the hardness of the urethane rubber tube is A50. Then, as shown in Fig.9, the urethane rubber tube expands
to the outside of the metal tube. As seen from Fig.8, the variation width of the pressure pulsation of the upstreamside of the component is about ±0.19MPa, and that of the downstream side of ±0.11MPa. Namely, the pressure
pulsation of the downstream side of the component becomes about 56% of that of the upstream side of it.
3.8

P2(MPa)P1(MPa)

3.8

3.6

3.6

3.4

3.4
3.2
0

20

40

60

80

100

3.2
0

20

40

60

80

100

Time(ms)

Time(ms)(a)Upstream(b)Downstream
Figure 8. Typical experimental results at 3.5MPa
Metal pipeUrethan rubber tube

Figure 9. Photos of urethan rubber tube at 3.5MPa

1.0

Amplitude (MPa)Amplitude (MPa)Figure 10 shows the results of frequency analysis for the experimental results in Fig.8. From Fig.10 (a), it is
known that there are peaks at almost the integral multiplication of frequency 225Hz. The relation between thefrequency and the amplitude are shown in Table 2. This table describes that the amplitude at 444Hz is 0.91MPain the upstream side, which is the largest and the amplitude in the downstream side at 444Hz is becomes about
41% smaller than the amplitude in the upstream side. Furthermore, it is known that the amplitudes in thedownstream side at 665Hz and 887Hz are about 28% smaller than the amplitudes in upstream side. As a result,
this component reduces the pressure pulsation by about 56% of the pressure pulsation of the upstream side of it.
Consequently, it is confirmed that this component works effectively to reduce pressure pulsation when the
average pressure at the pump discharge port is 3.5 MPa which is a practical pressure in a machine tool.

0.8
0.6

0.4

0.8
0.6

0.4
0.2

0.2
0
0

1.0

200

400

600

800

Frequency (Hz)

1000

0
0

200

(a)Upstream side

400

600

800

Frequency (Hz)1000(b)Downstream side
Figure 10. Frequency analysis resultsFrequency (Hz)
222
444
665
887

Table 2. Frequency and amplitude
Amplitude (upstream) (MPa) Amplitude (downstream) (MPa)
0.18
0.14
0.91
0.38
0.14
0.04
0.14
0.03

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Amplitude (MPa)In Figs 4 and 9, there is a larger peak at about twice the integral multiplication of frequency which is the productof its rotational frequency and its piston number. The reason is that the component to reduce pressure pulsation
is installed. To verify this point, the experiment was conducted when the component was replaced with astraight pipe in the experimental apparatus shown in Fig.3. Then, the average pressure at the pump discharge
port was set about 3.5MPa and the flow rate through the metering valve was set about 2.5×10-4m3/s (15L/min).
The result of frequency analysis for pressure P2 is shown in Fig.11. This figure shows that there is a large peakat the integral multiplication of frequency. Hence, it is known that the amplitude of twice the integral
multiplication of frequency becomes larger than the amplitude of the frequency due to the installation of thecomponent.
1.0

0.8
0.6

0.4
0.2
0
0

200

400

600

800

Frequency (Hz)1000

Figure 11. Frequency analysis results when the component is replaced with a straight pipe

Table 3 shows the experimental results when the gap width between the metal pipes and the hardness of theurethan rubber tube were varied. In this table, ΔP1 is the variation width of the pressure pulsation of the
upstream side of the component and ΔP2 that of its downstream side. As can be seen from this table, theperformance to reduce pressure pulsation is improved with an increase in the gap width between the metal pipes
and a decrease in the hardness of the urethan rubber tube. In this study, the performance of the component
becomes maximum in the combination of gap width 4.0mm and hardness A50. In this component, the gap width
can be increased up to 9.0 mm as shown in Fig.2. And the thickness and the internal diameter of the urethanrubber tube are important design parameters. Therefore, in next step, it is necessary to investigate the influenceof these parameters on the performance of the component to find suitable combination of the design parameters.
Table 3. Experimental results varying gap width and hardness
Gap width (mm)Hardness
ΔP1 (MPa)
ΔP2 (MPa)
ΔP2/ΔP1×100 (%)
2.0mm
A50
±0.185
±0.12
63
2.0mm
A70
±0.195
±0.14
71
4.0mm
A50
±0.190
±0.11
56
4.0mm
A70
±0.183±0.11
58

CONCLUSIONSIn this study, the structure of the improved component to reduce pressure pulsation was described. And theeffectiveness of the mechanism to reduce pressure pulsation and the performance of the component were madeclear through some experiments. Consequently, it is confirmed that the component can reduce the pressure
pulsation by 56% when gap width and hardness of the urethan rubber tube are suitable. Since there are otherdesign parameters, next step is to investigate their influence on the performance of the component. Especiallywhen the gap width is larger than 4mm, it seems to be expected that the performance of the component is
improved. In addition to it, it is necessary to develop the mathematical model of the component to determine thesuitable design parameters effectively.

ACKNOWLEDGMENTS
This work was supported by JSPS KAKENHI Grant Number 19K04152. The authors thank Mr. Ryuto Suzuki
who was a student in Ashikaga University and Mr. Taichi Sakato who is a student in Ashikaga University for
their cooperation with the experiments.

REFERENCES
1. ITO, Y., Thought-Evoking Approaches in Engineering Problems, Springer International Publishing, 2014, pp.167-168.

6

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The 11th JFPS International Symposium on Fluid Power 2020

2. LI, B., GUO, S., CAI, B. SHI, Z., Study on Pressure Pulsation Suppression of Reciprocating Pump, Journal of Physics:
Conf. Series, Vol.1187, Issue 3, 2019.
3. GHAZALI, M.H., HEE, L.M., LEONG, M.S., Piping Vibration due to Pressure Pulsations: Review, Advanced Materials
Research, Vol.845, 2014, pp.350-354.
4. ICHIYANAGI, T., KURIBAYASHI, T., NISHIUMI, T., Design Criteria of a Helmholtz Silencer with Multiple Degrees
of Freedom for Hydraulic Systems, Journal of Fluid Science and Technology, Vol. 8, No.3, 2013, pp. 277-293.
5. UMEDA, T., NISHIMURA, T., TSUTA, T., Fujiwara, M., Dynamic Response Analysis of a Tandem Type HydraulicPiston Pump with Swash Plate, Transaction of the Japan society of Mechanical Engineers, Part C, Vol.64, No.617, 1989,
pp.126-133. (in Japanese)
6. LOW, S. K, NAKANO, Y., KATO, H., Suppression of Chatter Marks due to Pulsation of Hydraulic Oil Pressure in
Surface Grinding, Journal of the Japan Society for Precision Engineering, Vol. 55, No.3, 1989, pp.514-519, 1989. (inJapanese)
7. LINIGER, J., SEPEHRI, N., SOLTANI, M., PEDERSEN, H. C., Signal-Based Gas Leakage Detection for Fluid Power
Accumulators in Wind Turbines, Energies, Vol.33, No.10, 2017, en10030372.
8. SAKURAI, Y., ETINOT, B., HYODO, N., AIBA, K., Proposal of a Component to Reduce Pressure Pulsation in Oilhydraulic System, Proc. of the 11th JFPS International Symposium, 2017.
9. HASHIMOTO, M., SAKURAI, Y., HYODO, N., AIBA, K., Proposal of Innovative Oil-hydraulic Component to Reduce
Pressure Pulsation, Proc of The 8th International Conference on Manufacturing, Machine Design and Tribology, 2019.
10. Practical hydraulic pocket book, Japan Fluid Power Association, 2021, pp.364. (in Japanese)7

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Genaral Session | Oil hydraulics

Construction, Components and Systems[GS6-01] Operating Information Presentation for Hydraulic Construction Robot〇Hironao Yamada1, Takahiro Ikeda1, Satoshi Ueki1, Kazuma Shinkai2, Katsutoshi Ootsubo3（1.Gifu University, 2.Sumitomo Heavy Industries, Ltd., 3.Kinjo Gakuin University）[GS6-02] Frequency Response Analysis of Parallel Link Mechanism using Oilhydraulic Cylinders of Tunnel Boring Machines
Ryo Yamamoto1, 〇Kazushi Sanada1, Shigeaki Ashikaga2 （1.Yokohama National University,
2.Komatsu Ltd.）[GS6-03] A Power-Split Hybrid Transmission to Drive Conventional Hydraulic Valve
Controlled Architectures in Off-road Vehicles: The Case of a MiniExcavator〇Mateus Bertolin1, Andrea Vacca1 （1.Purdue University）[GS6-04] Control of Air Bubble Content in Working Oil by Swirling Flow
Sayako Sakama2, 〇Yutaka Tanaka1, Yasuhiro Kodera3, Yoshiaki Kitamura3 （1.Hosei University,
2.National Institute of Advanced Industrial Science and Technology, 3.KYB Corporation）[GS6-05] Pulse Tests on Additive Manufactured Valve Blocks – Damage Analysis andNew Design Possibilities〇Sebastian Deuster1, Stefan Aengenheister1, Gunnar Matthiesen1, Katharina Schmitz1 （1.RWTH
Aachen University, Institute for Fluid Power Drives and Systems (ifas)）[GS6-06] The Potential in Fluid Power Systems for a Sustainable Future〇Katharina Schmitz1 （1.RWTH Aachen University, Institute for Fluidpower Drives and Systems(ifas)）[GS6-07] New Intelligent Hydraulic Power Control System〇Takahiro Urai1, Ken Shindo1 （1.Bosch Rexroth corporation）©The Japan Fluid Power System SocietyGS6-01

The 11th JFPS International Symposium on Fluid Power 2020

Operating Information Presentation for Hydraulic Construction Robot
Hironao YAMADA*, Takahiro IKEDA*, Satoshi Ueki*,
Kazuma SHINKAI**, Katsutoshi OOTSUBO ****Department of Mechanical Engineering, Faculty of Engineering, Gifu University
1-1 Yanagido, Gifu City, Gifu 501-1193, Japan(E-mail: yamada@gifu-u.ac.jp)**
Sumitomo Heavy Industries, Ltd.
ThinkPark Tower, 1-1 Osaki 2-chome, Shinagawa-ku, Tokyo 141-6025 Japan*** Department of Human Life and Environment, Kinjo Gakuin University
2-1723 Omori, Moriyama-ku, Nagoya, Aichi 463-8521 Japan

Abstract. Teleoperation system for construction robot is effective to work in hazardous environment. However, it isa problem to be solved that the work efficiency is inferior to the operation performed directly using a construction
machine. The purpose of this study is to improve the operation feeling for teleoperation. In this research, we havedeveloped a "Risk level warning system for avoiding approaching hazardous object" that warns the operator withforce and sound when the fork claw approaches a hazardous object. Furthermore, we conducted subject experiments
based on multiple force sense presentation methods, and clarified the optimal presentation method from theviewpoint of risk avoidance performance and worker burden.
Keywords: Construction Robot, Remote Control, Operating information, NASA-TLX

INTRODUCTION
Teleoperation system for construction robot is effective to work in hazardous environment [1][2]. However, it isa problem to be solved that the work efficiency is inferior to the operation performed directly using a
construction machine. In our research, we have been developing an intuitive teleoperation system for aconstruction robot in order to perform highly efficient and accurate work. This system is powered by hydraulicactuators and provides force feedback using a hydraulic pressure sensor attached to a cylinder. Recently, wedeveloped an AR presentation system that supports 3D recognition of the work field for an operator, andconfirmed its effectiveness for improving the work efficiency and accuracy [3], [4].
The purpose of this study is to improve the operation feeling for teleoperation. In particular, remote control
work is assumed in an environment where hazardous or fragile objects exist around the construction robot at a
disaster site. We will develop a "Risk level warning system for avoiding approaching hazardous object" thatwarns the operator by presenting force and sound when the fork claw approaches a hazardous object.
Furthermore, through subject experiments, we will examine the optimal presentation method from the viewpointof risk avoidance performance and worker burden.

Figure 1. Teleoperation system

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TELEOPERATED CONSTRUCTION ROBOT SYSTEM
The teleoperated construction robot system used in this research is shown in Figure 1. This system consists of
joysticks (JS), a construction machine, a video camera, a construction machine control computer (PC1), and an
image presentation computer (PC2). The master-slave control system is constructed by changing the hydrauliccylinder speed of the construction machine according to the input joystick displacement. The image of theremote place acquired by the video camera is presented to the PC2, and the operator watches the image andperforms the work.
The operator operates the joystick based on master-slave control. The displacement of the joystick is detected asa voltage signal by the potentiometer and input to PC1 via the A/D converter. Then, the calculated control input
is input to the servo amplifier as a voltage signal via the D/A converter. The servo amplifier converts the given
voltage signal into a current signal and drives the hydraulic servo solenoid valve. As a result, the piston operates,
and its displacement is detected by the potentiometer and input to PC1 via the A/D converter. The configuration
diagram of this system is shown in Figure 2. However, the parts not related to this study are omitted.
Construction Robot

Personal computer

Joystick

A/D
conv.
Amp.

Pump Unit

Figure 2. Schematic diagram of the hydraulic system

OPERATING INFORMATION PRESENTATION SYSTEM
In this system, the risk level is judged from the positional relationship between the fork claw's hand coordinates
Po, the contact judgment points Pa to Pd (the bottom point when the fork claw is a rectangular parallelepiped),
and the hazardous object. Then, the force sense and sound are presented according to the positional relationship(Figure 3, Figure 4). Here, these coordinates are based on the Cartesian coordinate system with the swing jointof the construction machine as the origin. In addition, it is assumed that the hazardous object is defined as arectangular parallelepiped shape composed of planes parallel to the coordinate system plane, and the coordinates
are known.
In addition, the space around the hazardous object that is expanded by 100 mm in each coordinate axis direction
is defined as the hazardous space, and the space is divided into 5 spaces in 20 mm increments as shown in
Figure 4. In this system, the risk level is defined in the reverse order of proximity to the hazardous object for thehazardous space and its external space. Then, the operator is presented with a sense of force according to therisk level and the distance to the hazardous object.

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Figure 3. Fork claw

Figure 4. Hazardous space (top view)

The risk level is L (an integer from 0 to 5), the distance to contact is D [mm], and the haptic value is F. The
force sense value calculated by any of the equations (A) to (D) is presented in the direction opposite to the input
for all axes of the joystick operation. When the risk level L is 0, the distance D to contact is fixed at 100 mm.

F  0. 1  L(A)
2

F  0.03  L
F  0.005  (100  D)

(B)
(C)

F  0.000075  (100  D) 2

(D)

Figure 5. Haptic presentation pattern
Presentation patterns A and B using the force sensory equations (A) and (B) are presented step by step
according to the risk level consisting of 6 levels. On the other hand, the presentation patterns C and D using
equations (C) and (D) perform stepless presentation. In addition, patterns B and D, which are the squares of eachvariable, change based on the exponential function. On the other hand, patterns A and D change based on linear
functions.

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The haptic values calculated by the following patterns (A) to (D) as shown in Figure 5 are presented to thecontrol lever. In addition, a warning sound is presented at the presentation cycle T [s] calculated by equation (E).
When presenting a warning sound, a wav file with a length of about 0.2s is read.

T  (6  L ) / 5(E)

EXPERIMENTAs an evaluation experiment of this system, a subject experiment was conducted for the gripping work of an
object. The work environment is shown in Figure 6. In the work environment, three cardboard boxes assuming
hazardous objects, a wooden block that is the gripping object to be transported, and one wooden stand that is thetransporting point of the gripping object are installed. Figure 7 shows a layout diagram of the work field. Six
subjects were asked to carry the wooden block 2 to the transport point 3 so as not to touch the hazardous object
1. The subject works while watching a video showing the work environment. The experiment is performed by
five presentation methods including E: sound presentation in addition to four types of force sense presentation
based on the calculation methods A to D. The subjects performed the prescribed operations twice for eachpattern, and measured the work time Tf [s], risk index R, and mental burden index. In addition, the evaluation
was performed by a questionnaire after the experiment. Here, when the risk level at each time is Lt, the risk
index R is expressed by Eq. (1), which is the sum of the risk levels divided by the working time.
TfR   Lt / T f

(1)

t 0

Figure 6. Work field

Figure 7. Work field (Top view)

In the experiment, sufficient practice is performed in advance to reduce the influence of proficiency on theexperimental results. The subjects were 5 males and 1 female aged 21 to 25 years, with an average age ofapproximately 22.8 years. Four of them have experience in operating construction machinery, and the remaining
two have no experience in operating construction machinery. In addition, considering the order effect, theexperiments were conducted in a different order for all the subjects.

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EXPERIMENTAL RESULTS AND DISCUSSION
Figure 8 shows the results of the subject average of the risk index based on the results of the experiment. From
the figure, it can be seen that the risk index of B is the lowest. In addition, when multiple comparisons were
performed by the Turkey test (single-step multiple comparison procedure and statistical test), a significant
difference (*) with a significance level of 5 [%] was observed between patterns A-B, B-C, and B-E. Therefore,
the presentation method using pattern B is considered to be optimal from the viewpoint of risk avoidance.
Although no significant difference was confirmed, the risk index was also lower in pattern D than in patterns A,
C, and E. This suggests that the method of increasing the force sense value exponentially is suitable forpresentation.

Figure 8. Average of risk index
In general, the magnitude of human sensation increases in proportion to the logarithm of the intensity of thestimulus [1]. Therefore, even when a large force sense value is presented, it is easy to intuitively grasp thechange in force sense due to approaching a hazardous object. It is considered that this prevented further
approach and contact. In addition, although no significant difference was confirmed, the risk index of pattern B,
which changes stepwise, was lower than that of pattern D, which changes the force sense value steplessly.
Discrete changes cause the force sense value to rise suddenly at the moment the risk level rises. Therefore, theoperator could intuitively grasp that the fork claw was approaching a hazardous object. Therefore, it isconsidered that the risk index was suppressed. In this study, we used “NASA-TLX” [5] for the assessment of themental task load. The index is given subjectively, and the task load is divided into six indices: Mental Demand(MD), Physical Demand (PD), Temporal Demand (TD), Performance (OP), Effort (EF) and Frustration (FR).
The larger values correspond to larger mental task loads.

Figure 9. Average of NASA-TLX

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Figure 9 shows the result of averaging the six evaluation indexes of NASA-TLX. From the results of Figure 9, it
can be seen that the overall mental burden increases in the cases of patterns A and E. In the presentation method
using pattern A, the force sense value changes rapidly when the risk level rises. For this reason, it is likely thatthe work will be burdensome at low risk levels. Similarly, in the case of pattern E, it is considered that it was aburden on the operator that the warning sound was constantly emitted when trying to grasp the grasped object
surrounded by hazardous objects. On the other hand, the results showed that the overall mental burden was
lower in patterns B, C, and D. Comparing the patterns C and D in which the force sense value was increased
steplessly from the results, no significant difference could be confirmed in MD, PD, OP, and EF. However, in
FR, pattern D had a slightly higher mental burden than pattern C. This is thought to be because pattern D has ahigher maximum force sense value. It can be said that the larger the force sense value, the more stress theworker feels. In addition, pattern B, in which the force sense value is gradually increased, has a higher degree ofmental burden in PD and FR than in patterns C and D. However, the mental burden was lower in MD, OP, andEF. When the risk level rises, the sudden rise in the force sense value becomes a physical burden and stresses
the worker. However, since the operation information is easy to understand, it is easy to be satisfied with thework results and the mental burden is considered to be low.

CONCLUSION
In this research, we have developed a "Risk level warning system" that warns the operator with force and soundwhen the fork claw approaches a hazardous object. Furthermore, we conducted subject experiments based on
multiple force sense presentation methods, and clarified the optimal presentation method from the viewpoint ofrisk avoidance performance and worker burden. In the experiment, the blocks were transported by 6 subjects
twice for each pattern, for a total of 10 times. As a result, among the proposed methods, the stepwise
presentation pattern based on the exponential function is considered to be the most effective presentation
method comprehensively.

REFERENCES
1. J. Xue, L. Zhang, T. E. Grift, “Variable field-of-view machine vision based row guidance of an agricultural robot”, inJournal of Computers and Electronics in Agriculture, 84, 2012, pp. 85–91.
2. C. L. Perkinson, M. E. Bayraktar, I. Ahmad, “The use of computing technology in highway construction as a total jobsite
management tool”, in Journal of Automation in Construction, 19(7), 2010, pp. 884–897.
3. Support System for Slope Shaping Based on a Teleoperated Construction Robot, Katsutoshi Ootsubo, Daichi Kato,
Takuya Kawamura and Hironao Yamada, Journal of Robotics and Mechatronics, Vol. 28，No. 2, pp.149-157, 2016.
4. Teleoperated Construction Robot Using Visual Support with Drones, Hironao Yamada, Naoyuki Bando, Katsutoshi
Ootsubo, and Yuji Hattori, Journal of Robotics and Mechatronics, Vol. 30，No. 3, pp.406-415, 2018.
5. S. Hart, L. Staveland, “Development of NASA-TLX (Task Load Index): Results of empirical and theoretical research”,
in Advances in Psychology, Human mental workload, 52, North Holland, 1988, pp. 139–183.

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Frequency Response Analysis of Parallel Link Mechanism
using Oil-hydraulic Cylinders of Tunnel Boring Machines
Ryo YAMAMOTO*, Kazushi SANADA**, and Shigeaki ASHIKAGA****

Department of Mechanical Engineering, Materials Science, and Ocean Engineering,
Graduate School of Engineering Science, Yokohama National University**
Faculty of Engineering, Yokohama National University
79-5, Tokiwadai, Hodogaya Yokohama city, Kanagawa, 240-8501 Japan(E-mail: sanada-kazushi-sn@ynu.ac.jp)***
Komatsu Ltd.

Abstract. Frequency responses of a parallel link mechanism using six oil-hydraulic cylinders of tunnel boring machine
are analyzed in this study. Each oil-hydraulic cylinder is driven with a constant-displacement oil-hydraulic pumpwhose rotational frequency is controlled by an AC servo motor. The six cylinders generate force for digging the ground.
A mathematical model of the parallel link mechanism of six oil-hydraulic cylinders is used in this study. Statics of theparallel link is described with Jacobian matrix. A dynamic model of the oil-hydraulic cylinder and oil-hydraulic pumps
is built based on the equation of motion of cutter head, compressibility of oil, and pump characteristics. A model ofthe tunnel boring machine is obtained by combining the models of parallel link and oil-hydraulic circuit. Frequency
responses of the model are useful to investigate control methods of the tunnel boring machine.
Keywords: Cylinder, Parallel link, Frequency response, Mathematical model, Oil-hydraulics

INTRODUCTION
Tunnel boring machines are heavy-duty machines which are used for boring tunnels. An important recent trend
in research and development is automatic control of tunnel boring machine. The tunnel boring machine is required
to follow planed path of tunnel. Accuracy of tunnel boring is influenced by various factors, for examples, nonuniform hardness of bedrock and machine performance. The tunnel machine treated in this study consists of aparallel link mechanism of 6 oil-hydraulic cylinders driven by pump control. Statics of parallel link was studied
by ARAI [1]. Link lengths, forces, and torques are related each other by Jacobian matrix. TSUCHIYA, et al.
investigated a parallel link mechanism with 6 oil-hydraulic cylinders controlled by servo valves [2][3]. Combining
statics of the parallel link mechanism and dynamics of the servo-valve controlled cylinder drive system, resultsof frequency responses were compared with experimental results. A simple inertia load was considered in thestudies. When the tunnel boring machine is investigated, various forces and torques due to non-uniform hardnessof bedrock are necessary to be considered. Path control of cutter head is a main task for automatic control system.
In addition, instead of servo-valve controlled cylinder drive system, one-pump one-cylinder drive system isapplied. A fixed-displacement oil-hydraulic pump, whose rotational frequency is determined by an AC
servomotor, directly drives one cylinder. Final goal of this project is to develop automatic control system for thetunnel boring machine to follow pre-specified path. In this study, as the first step, modeling and frequency
response analysis are investigated.
NOMENCLATURES
𝐴
𝐵
𝑩

𝐵

𝑪
𝐷
e
𝑬

𝑥, 𝑦, 𝑧, 𝛼, 𝛽, 𝛾

𝑬

𝐸

, 𝐵

, 𝐸

, 𝐵

, 𝐸

𝑭
𝐹 ,𝐹 ,𝐹 ,𝑇 ,𝑇 ,𝑇
𝐹 ,𝐹 ,𝐹𝒇
𝑓 ,𝑓 ,⋯,𝑓area of piston bottom
viscous damping coefficient
position vector of m-th joint of base platform with respect to the origin ofbase platform
position and attitude angle vector of cutter headpump displacement
position vector of m-th joint of end effector with respect to the origin of end
effectorposition vector of m-th joint of end effector with respect to the origin of thebase platform
force and torque vector of cutter head
forces acting on the center of gravity of cutter head
cylinder force vector
1

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𝐽 ,𝐽 ,𝐽𝑱
𝑱
𝐾𝑙𝒍𝑙 ,𝑙 ,⋯,𝑙𝑀
𝑴 𝑑𝑖𝑎𝑔 𝑀, 𝑀, 𝑀, 𝐽 , 𝐽 , 𝐽𝒑
𝑝 ,𝑝 ,⋯,𝑝𝒑𝒒
𝑞 ,𝑞 ,⋯,𝑞𝑹𝑹𝛼, 𝛽, 𝛾𝑹𝑠
𝑇 ,𝑇 ,𝑇𝑡
𝑉𝑥, 𝑦, 𝑧𝑥 ,𝑦 ,𝑧𝛼, 𝛽, 𝛾𝜂∗

function relating 𝑪 and 𝒍moment of inertia of cutter head
Jacobian matrix
coefficient matrix
bulk modulus of oil
length of m-th link
vector of link length
mass of cutter headmatrix of mass and moment of inertia of cutter head
position vector of the origin of end effector with respect to the origin of baseplatformbottom pressure vector
pump flow rate vector
demand signal of position and attitude angle vector 𝑪
demand signal of link length vector 𝒍rotation matrix
Laplace variable
Torques acting on the center of gravity of cutter headtime
oil volume of bottom side of cylinder and pipe/hosecoordinate system of base platform (gripper)
coordinate system of end effector (cutter head)
attitude angle of cutter head
volumetric efficiency
operating point
small variation
TUNNEL BORING MACHINE

A tunnel boring machine which is investigated in this study is shown in Figure 1. It consists of a cutter head, 6cylinders, and a gripper. The cutter head is rotated by an oil-hydraulic motor. Disk cutters of the rotating cutter
head break bedrock, and tunnel boring is performed. The cutter head is pushed by a parallel link consisting of 6cylinders. The gripper maintains machine position inside a tunnel. Roof support, side support, and vertical support
are used to cover the tunnel wall and support a body of the tunnel boring machine. One-pump one-cylinder drive
system is applied to drive the 6 cylinders. Control of the parallel link determines direction of tunnel which is dugby the tunnel boring machine. The aim of this study is research and development of automatic control system forthe tunnel boring machine. Because the tunnel boring machine consists of the parallel link of 6 cylinders and theoil-hydraulic circuit of one-pump one-cylinder drive system, a mathematical model is built by combining eachsub-system in the next chapter.

Figure 1. Tunnel boring machine: Six cylinders push a rotating cutter head to break bedrock in the Z-direction.

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MATHEMATICAL MODEL
Parallel Link Mechanism
The parallel link mechanism consists of end effector (cutter head), base platform (gripper), and 6 links (cylinder
No.1, 2, …, 6) as shown in Figure 2. Coordinate system 𝑥𝑦𝑧 and 𝑥 𝑦 𝑧 are fixed on base platform and end
effector, respectively. Position of the origin of the end effector 𝑥 𝑦 𝑧 with respect to the origin of the base𝑥, 𝑦, 𝑧 . Six broken lines represent six cylinders. 0𝑩 andplatform 𝑥𝑦𝑧 is presented by a position vector 𝒑e
𝑬 are position vectors of cylinder joints of the base platform and the end effector, respectively.

Figure 2. Parallel link consisting of end effector, base platform, and 6 links

Equation of Motion of Cutter Head
Cutter head motion is a combination of transrational and rotational motions. The equation of motion of the cutter
head is written as:
𝐹

𝑀𝑥,

𝐹

𝑀𝑦,

𝐹

𝑀𝑧

𝑇

𝐽 𝛼, 𝑇

𝐽 𝛽,

𝑇

𝐽𝛾

,

(1)

where position of the center of gravity of the cutter head 𝑥, 𝑦, 𝑧, and attitude angle of the cutter head 𝛼, 𝛽, 𝛾. A
vector 𝑪 is used to represent the variables:𝑪𝑥, 𝑦, 𝑧, 𝛼, 𝛽, 𝛾.

(2)Forces and moments are written by a vector 𝑭:
𝑭

𝐹 ,𝐹 ,𝐹 ,𝑇 ,𝑇 ,𝑇.

(3)

The equation of motion can be rewritten in a vector form:
𝑭

𝑴𝑪 .

(4)

Statics of Parallel Link Mechanism
Statics of parallel link [2][3] was applied to build a mathematical model of the tunnel boring machine. The
position vector of the m-th joint of the end effector with respect to the origin of the base platform, 0𝑬 , is writtenas:3

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𝟎

𝑬𝒎

𝑥 𝑦 𝑧

𝛼, 𝛽, 𝛾 𝒆𝑬𝒎 ,𝑹

where a rotation matrix of Euler angle of the end effector, 𝑹𝑹

cos𝛽cos𝛾
sin𝛼sin𝛽cos𝛾 cos𝛼sin𝛾
cos𝛼sin𝛽cos𝛾 sin𝛼sin𝛾

𝛼, 𝛽, 𝛾(5)𝛼, 𝛽, 𝛾 , is written as:

cos𝛽sin𝛾
sin𝛼sin𝛽sin𝛾 cos𝛼cos𝛾
cos𝛼sin𝛽sin𝛾 sin𝛼cos𝛾

sin𝛽
sin𝛼cos𝛽 .
cos𝛼cos𝛽(6)

Joint distance of m-th link, 𝑙 , is calculated by:
𝐸

𝑙

𝐵

𝐸

𝐵

𝐸

𝐵

.

(7)

Link length vector 𝒍 is a function of a vector 𝑪 :
𝒍

ℎ 𝑪 .

(8)

Assuming an operating point (denoted by *), small variation (denoted byconsidered, then:
𝒍∗

𝜕ℎ 𝑪
𝜕𝑪

𝒍 ≅ ℎ 𝑪∗) about the operating point is

𝑪.

(9)

𝑪 𝑪∗

The partial-differentiation coefficient is defined as Jacobian matrix 𝑱 which is a 6
𝜕ℎ 𝑪
𝜕𝑪

𝑱

.

𝑪 𝑪∗

6 matrix:(10)

Small variation of link length is calculated from small variation of the position and attitude angle vector multipliedby the Jacobian matrix:𝒍

𝑱 𝑪∗ 𝑪 .

(11)

Small variation of demand signal of link length is calculated from small variation of the position and attitude angle
vector multiplied by the Jacobian matrix:
𝑹𝒍

𝑱 𝑪∗ 𝑹𝒄 .

(12)

According to the virtual work principle, a force and torque vector of the end effector (cutter head) 𝑭 is relatedwith a cylinder force vector 𝒇 using a transposed matrix of Jacobian matrix, 𝑱 𝑪 , then:𝑭

𝑱 𝑪 𝒇.

(13)

Linearizing the above equation:𝑭∗𝑭 ≅ 𝑱 𝑪∗ 𝒇∗

𝜕 𝑱 𝑪 𝒇
𝜕𝒇

𝒇 𝒇∗
𝑪 𝑪∗

𝒇

𝜕 𝑱 𝑪 𝒇
𝜕𝑪

𝒇 𝒇∗
𝑪 𝑪∗

𝑪.

(14)

Small variation 𝑭 is written as:
𝑭

𝑱 𝑪∗ 𝒇

𝑱 𝑪∗ 𝑪 ,(15)

where a coefficient matrix of the second term in the right-hand side is𝑱 𝑪∗

𝜕 𝑱 𝑪 𝒇
𝜕𝑪

𝒇 𝒇∗
𝑪 𝑪∗

.

(16)

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Oil-hydraulic Circuit
Schematic diagram of one-pump one-cylinder drive system is shown in Figure 3. The tunnel boring machine of
our interest has six sets of the one-pump one-cylinder drive system for each link of the parallel link mechanism
shown in Figure 2. A cylinder (a) is driven by a pump (b) whose rotational frequency is determined by an AC
servo motor (c). Maximum outlet pressure of pump is limited by relief valves (d). When the cylinder extracting,suction flowrate of pump is shortage, and a check valve (e-1) will open to fulfil the suction oil flow of the pump(a). When the cylinder retracting, the cylinder receives force from bedrock, and bottom pressure 𝑝 is generated
with a counter-balance valve (f-2). Pump suction flowrate from the bottom side is larger than flowrate to the rod
side of the cylinder. An excess of the suction flowrate over the rod-side flowrate will relieve through a shuttle
valve (g) to tank. The m-th cylinder generates cylinder force 𝑓 . A cylinder force vector is written as:𝒇

𝑓 ,𝑓 ,⋯𝑓.

(17)

The cylinder force is equal to the amount of pressure force minus viscous damping force:
𝒇

𝐴 𝒑𝟐

𝐵

𝒍 ,

(18)

where piston area 𝐴 and viscous coefficient 𝐵 on the assumption that 6 cylinders have the same piston. Bottompressure vector is 𝒑 , and link length vector is 𝒍. Linearization of this equation and substitution of Eq.(11) for
link length result in:
𝒇

𝐴 𝒑𝟐

𝐵

𝑑
𝒍
𝑑𝑡

𝐴 𝒑𝟐

𝐵

𝑑
𝑱 𝑪∗ 𝑪 .
𝑑𝑡(19)

A pump flowrate vector of the hydraulic pump, 𝒒𝟐 , is determined by a rotational frequency vector, 𝝎. The pumphas displacement, 𝐷, and volumetric efficiency, 𝜂 :𝒒𝟐

𝜂 𝐷𝝎 ,(20)𝒒

𝜂 𝐷𝝎 .

(21)

Linearizing this equation:

Bottom pressure is determined by pump flowrate and piston velocity according to compressibility of oil:
𝑑
𝒑
𝑑𝑡 𝟐

𝐾
𝒒𝟐
𝑉

𝐴

𝑑
𝒍 .
𝑑𝑡

(22)

Linearizing the above equations and substituting Eq.(11) for 𝒍 :
𝑑
𝒑
𝑑𝑡

𝐾
𝒒
𝑉

𝐴

𝑑
𝒍
𝑑𝑡

𝐾
𝒒
𝑉

𝐴

𝑑
𝑱 𝑪∗ 𝑪 .
𝑑𝑡

(23)

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Figure 3. One-pump one-cylinder drive system

Block Diagram of Tunnel Boring Machine
A block diagram of the tunnel boring machine is drawn in Figure 4, which are obtained based on the linearizationof fundamental equations. Pump flowrate 𝒒 is determined by rotational frequency 𝝎, pump displacement 𝐷, andvolumetric efficiency 𝜂 , Eq.(21). Bottom pressure 𝒑 changes based on the amount of pump flowrate 𝒒 minus
piston-velocity induced flowrate 𝐴 𝑱 𝑪∗ 𝑠𝑪 according to compressibility of oil, Eq.(23). Cylinder force 𝒇 isequal to the amount of pressure force 𝐴 𝒑 minus viscous damping force 𝐵𝑱 𝑪∗ 𝑠𝑪, Eq.(19). Force and torqueacting on the cutter head is the sum of cylinder force component 𝑱 𝑪∗ 𝒇 and link mechanism component 𝑱 𝑪∗ 𝑪,
Eq.(15). According to the equation of motion of cutter head, Eq.(4), the vector of displacement and attitude angleof the cutter head, 𝑪, is derived. Finally, link length 𝒍 is calculated from the vector 𝑪, Eq.(11).

Figure 4. Block diagram of linearized model of the tunnel boring machine

Frequency Response Analysis
Frequency responses from demand signal of rotational frequency of AC servo motor to link length are calculatedusing the block diagram of Figure 4. The results are shown in Figure 5. Horizontal axis is frequency from 0.01 to
10 Hz. Upper plot is gain plot in dB, and lower plot is phase plot in degree. The system has six demand signals ofrotational frequency as input signal and six displacements of cylinder as output signal. Considering all
combinations from 6 input signals to 6 output signals, the number of transfer functions becomes 36. In Figure 5,
6

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only 6 sets of frequency responses from an input signal of one of AC servo motors of the cylinder No.1 to output
signals of 6 cylinders are plotted in Figure 5. The solid line “from 1 to 1” means autocorrelation, which isfrequency response of a demand signal of rotational frequency of AC servo motor to displacement of cylinder
which is directly connected with a pump driven by the AC servo motor whose rotational frequency is taken as the
input signal. In lower frequency region less than 0.4 Hz, the gain plot shows a slope of -20dB/decade and phaseof -90 degree. In higher frequency region, the phase decreases gradually to -180 degree. It can be suggested thatan integral element is included in the transfer function. Other lines “from 1 to 2”, “from 1 to 3”, “from 1 to 4”,“from 1 to 5”, and “from 1 to 6” are frequency responses of transfer functions representing mutual interference
from cylinder No.1 to other cylinders, No.2, No.3, No.4, No.5, and No.6, respectively. In lower frequency region
less than 0.4Hz, the gains are much smaller than gain of “from 1 to 1”. In higher frequency region, the differences
become relatively small. Phases of the mutual interference begin as 90 degree and 270 degree in lower frequency
region. It can be seen when considering complex plane, the “90” degree and “270” degree are identical to the 270 degree and -90 degree, respectively. Phase decrements of the mutual interferences are larger than that of“from 1 to 1”.
It can be said that bandwidth of the system is almost 0.4 Hz. In frequency region less than 0.4 Hz, a transfer
function of “from 1 to 1” can be approximated by an integral element, 1/𝑠 . Then, proportional control strategy
can be applied to feedback control of link length. When designing control system of the parallel link mechanism,
it is desired to consider the mutual interference among cylinders.
0Gain (dB)from 1 to 1
from 1 to 2
from 1 to 3

from 1 to 4
from 1 to 5
from 1 to 6

-100

Phase (deg.)

-200
360
180
0
-180
-360
-2
10

10

-1

10

0

10

1

Frequency (Hz)Figure 5. Frequency responses of a cylinder from rotational frequency of pump to length of link

CONCLUSION
Frequency response was calculated using a block diagram of mathematical model considering statics of theparallel link mechanism and the one-pump one-cylinder drive system. The main transfer function of interest canbe approximated as an integral element in low frequency region of less than 0.4Hz. Proportional control can be a
candidate for control strategy. A possible future work is feedback control of link length for the tunnel boringmachine.
REFERENCES
1. ARAI, T., Analysis and Synthesis of a Parallel Link Manipulator Based on Its Statics, Journal of the Robotics Society ofJapan, 1992, 10-4, pp.526-533.
2. TSUCHIYA, K. et al., Dynamic Characteristics of Parallel Link Mechanism with Six Degrees of Freedom Using Hydraulic
Cylinders, 1st Report Frequency Response Analysis of a Servomechanism, Transactions of the Japan Fluid Power System
Society, 2006, 37-1, pp.1-7.
3. TSUCHIYA, K. et al., Dynamic Characteristics of Parallel Link Mechanism with Six Degrees of Freedom Using Hydraulic
Cylinders, 2nd Report: Characteristic Analysis by Nonlinear Simulation and Experiment, Transactions of the Japan Fluid
Power System Society, 2008, 39-1, pp.1-6.

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A Power-Split Hybrid Transmission to Drive Conventional Hydraulic
Valve Controlled Architectures in Off-road Vehicles: The Case of a
Mini-Excavator
Mateus BERTOLIN* and Andrea VACCA**

Maha Fluid Power Research Center, Purdue University
1500 Kepner Dr., Lafayette, IN, United States(mbertoli@purdue.edu and avacca@purdue.edu)

Abstract. In recent years, several disruptive solutions for off-road vehicles hydraulic implements were proposed toincrease energy efficiency, many focusing on the reduction of throttling losses. However, the market has showed alimited adoption of such solutions due to high implementation costs and operator’s acceptance. Considering theselimitations, this paper proposes to adopt a hybrid power-split hybrid transmission between the supply pump and theprime mover, while still using state-of-the-art hydraulic actuation. Such approach results in an independent control ofthe engine and pump speeds, therefore improving the efficiency of both. The paper describes the potential applicationof the proposed solution to a small excavator equipped with a load sensing system. The authors developed a simulation
model, which was validated against experiments of the baseline machine. The model was used evaluate fuel efficiency
gains achieved with the proposed solution, which are estimated in up to 16% for a truck-loading cycle.
Keywords: Hybrid, excavator, power-split transmission, load-sensing system.

INTRODUCTION
Typical hydraulic architectures commonly found in off-road vehicles are well known for two main characteristics:
precise motion control, and low energy efficiency associated with the inherent throttling losses of their valvecontrolled systems. Especially in machines with multiple simultaneous actuations like excavators, losses causedby the valves are inevitable when a centralized hydraulic supply is used. This is because each pump in the machine
usually supplies more than one actuator at the same time. In addition, many vehicles today available in market
operate in not optimal conditions for the engine and the pumps. This happens because the faster dynamics and thehigh installed power capacity of the hydraulic pumps force the engine to operate at high speeds to avoid stall.
Additionally, the power and torque limitation of the engine usually results in partial displacement operation of thehydraulic pumps, leading to lower volumetric efficiency. Consequently, during high-power demanding cycles thesystem operates with two of its most important components at low efficiency conditions.
Literature has been consistent with two approaches to improve fuel efficiency of off-road vehicles. When possible,
hybridization is a good solution to improve engine efficiency and to reduce the maximum engine power
requirement. In addition, a lot of focus has also been given to minimization or elimination of throttling losses withsolutions such as independent metering and de-centralized displacement-controlled (DC) architectures.
Among the most successful systems proposed in literature are the ones that can achieve both a better engine
management and a reduction of throttling losses. Vukovic, Sgro, and Murrenhoff [1], [2] proposed the STEAM
excavator, which uses two accumulators to maintain three different pressure levels in the machine. The concept
is known as constant pressure rails, even though in reality the pressures can vary considerably in each accumulator
during the machine operation. On/off valves are used to connected different pressure levels to each actuatorchamber. A controller was developed to select the proper pressure level and chamber connection with the goal ofminimizing throttling losses and the actuator motion is controlled with proportional valves. The large
accumulators adopted in the STEAM allow meeting high flow demands decoupling the actuators flow requestfrom the engine speed. A 55% increase in efficiency during an air grading cycle was reported in the mentioned
work.
In a similar fashion, Heybroek and Sahlman [3] also propose different pressure rails with large accumulators tominimize throttling losses and decouple flow request from engine speed. In their proposal, two pressure levels areused in combination with multi-chamber cylinder with 4 chambers, achieving up to 35% reduction in fuelconsumption for a truck-loading cycle.
The author’s research lab has followed the de-centralized approach to eliminate completely throttling losses. It
has been shown by Zimmermann, Busquets and Ivantysynova that only by replacing the state-of-the-art hydraulics
with displacement control actuation it is possible to reduce fuel consumption by 40% [4]. Later, the machine was1

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modified to a series-parallel hybrid solution and achieved extra 10% in fuel savings in simulation [5], [6].
However, this solution keeps the DC pumps mechanically connected to the engine shaft. In this scenario, whenever
the engine speed is kept constant and the flow request is low, the pumps are forced to operate at low displacements
where the volumetric efficiency is low. Williamson [7] showed that by varying the engine speed this problem canbe solved, but nevertheless, the engine is forced to constantly vary its speed moving away from its optimal
operating point during high flow-demand. To address this issue, a power-split hydraulic hybrid transmission
between engine and hydraulic pumps was first proposed by the authors in a recent work [8]. With this approach,DC pumps and engine speeds can be controlled independently, leading to fuel consumption improvements of up
to 60% when compared to the baseline Load-Sensing system or extra 11% when compared to the previous versionof the hybrid system.
Despite the reduction in fuel consumption achieved in the past mentioned works, it must be considered that themarket has showed a limited acceptance of DC based solutions. Besides implementation costs, there are challenges
associated with reproducing all the machine requirements and expert operator feelings that are achieved by thecurrent state-of-the art hydraulic architectures, such as load-sensing and open-center systems.
To overcome these limitations, this paper proposes a solution that can be considered as less disruptive whencompared to the above-mentioned work, but that is able to achieve reduction in fuel consumption withoutcompromising the current operator experience and market acceptance of the hydraulic system architecture. The
rationale is to design a system with high engine-hydraulics integration with limited implementation costs. In
addition, the proposed solution aims at keeping the same motion control quality and user feeling of the state-ofthe-art hydraulic implements.
The paper structure is as follows. In the next section, the reference machine is introduced and followed by adescription of the proposed solution. Then, the modeling approach used to model the baseline load-sensing system
is described. Next, the reference working cycle used in this study is presented and simulation results are validated
against measurements. Following the model validation, the design of the proposed transmission is discussed, andthe simulation approach used to model the transmission is presented. Afterwards, the proposed powermanagement controller is presented, followed by a discussion on the possible integration of the swing actuator in
the transmission. Lastly, simulation results are presented and discussed, followed by conclusions.

REFERENCE MACHINE
The reference machine used in this study is a Bobcat 435 mini-excavator, which in its baseline version uses a
post-compensated load-sensing (LS) system with a single pump, like the one shown in Figure 1. The main machineparameters are shown in Table 1.
Table 1. Reference Machine Main Parameters

Machine Parameter
Engine PowerOperating Weight
Engine Operating Speed (High Speed Mode)
LS Pump Displacement
LS Pump Margin
LS Compensator Margin

Symbol

Value

𝑃!
-

𝑠

36 kW
5 tons.
2400 rpm
48 𝑐𝑚# /rev
23 bar

𝑠$

4 bar

𝑛!
𝑉"

In this architecture, the highest user pressure is picked-up by the LS line through each actuator pressurecompensator. The pump adjusts its displacement to keep its outlet pressure at the level of the LS line plus a fixedmargin (𝑠). Pressure compensators ensure a constant pressure differential across each flow control valve. Thispressure differential is equal to 𝑠 − 𝑠! and ensures that the flow is directly proportional to the valve opening area,
therefore eliminating any load interference on actuator speed. A fixed orifice bleeds out any trapped pressure in
the LS line, while a relief valve in the same line limits the pump maximum pressure. Pilot lines and local actuator
relief valves are omitted in Figure 1 for clarity.

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Boom

Arm

BucketSwing

Figure 1. Reference Machine and its baseline hydraulic schematic.

PROPOSED SOLUTION
The proposed solution consists in using a power-split transmission between the engine and the load-sensing pump,like shown in Figure 2.The idea is to improve the overall system fuel consumption, without major changes to thehydraulic actuation of the baseline machine. In addition, another solution including a hybrid swing integrated in
the transmission hydraulic path using secondary control will also be introduced and analyzed later in this paper.
Despite the proposed solution can be applied with any centralized fluid power architecture, this work demonstrates
its application on an excavator equipped with a LS system. It should be noted that a LS system is used here
exclusively because it was the original configuration of the baseline machine, for which data was available.
However, implementation with an open-center system is also possible.
The widely known power-split transmission has the characteristic of a series hybrid, in which engine and pumpspeeds are completely decoupled, but with a higher efficiency, since most of the power is still transmitted
mechanically through the planetary gear train (PGT). Another advantage is on the size of the transmission units
needed. While a series hybrid transmission would have both units sized for maximum system power, the selected
power-split system requires smaller units since it relies on mechanical power transmission as well.
The planetary gear used has three shafts. One can be connected to the engine, one the primary hydraulic unit andanother connected to the main user, in this case is a LS pump.

Transmission Power Flow
At the PGT, the torque relationship between the shafts is constant. On the other hand, the speed relationshipbetween the three shafts is given by𝑛"#$ − 𝑖% 𝑛&'$( − (1 − 𝑖% )𝑛!)&&'*& = 0(1)

where 𝑖% is the PGT standing gear ratio. Therefore, The PGT gear configuration enables independent speed controlof 2 out of 3 shafts, as clear from the expression stated in equation (1). Consequently, power can be split between
two paths according to the ratio of the selected speeds. In this way, while part of the engine power is transmitted
to the LS pump through the mechanical path, the remaining power is transmitted through the hydrostatic path of3

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the transmission. The ratio between inlet and outlet flows in the transmission hydraulic path determines theaccumulator state-of-charge, allowing for energy storage/release.
With this approach, the primary unit is used to load the engine with the desired torque, while the secondary unit
uses secondary control to regulate the speed of the LS pump. The engine speed is then controlled by adjusting theengine throttle, with a setpoint determined by a power-management controller. Therefore, compared to the original
system the engine is operated in a much steadier way, with nearly constant torque and speed, while theaccumulator and the secondary unit handle the severe load transients that occur during a digging operation. In
addition, the output shaft speed can be quickly adjusted over time such that the LS pump is forced to continuously
operate close to its maximum displacement, hence increasing its volumetric efficiency. Additionally, when themechanically transmitted power is higher than the demand from the LS pump, the secondary unit serves ashydrostatic brake, also charging the accumulator. Additional relevant information on the behavior of hydraulicpower-split transmissions can be found in [9] and [10].
Besides the freedom in engine speed selection, the proposed architecture also brings the typical benefits of hybrid
systems, such as the potential for engine downsize. The added energy storage device allows the designer to use asmaller engine such that it is constantly operated closer to its maximum torque, where engine efficiency ismaximum, while peak power demand is achieved with a hydraulic boost from the hydraulic accumulator.

Boom

Arm

BucketSwing

Pe

Transmission

Pmech Mechanical Path

Phyd

PLS

Pacc
P2

Transmission
Hydrostatic Path
Primary Unit

Secondary Unit

Figure 2. Proposed power-split hybrid excavator with load-sensing implements.

SIMULATION MODEL
The design of the proposed transmission is based on a numerical simulation model that is able to reproduce thebehavior of the baseline LS system and achieve a good estimate of its power consumption. Such model is used tocharacterize the loads that would be seen by a power-split transmission, and to provide important information in
the design process. In addition, it is also used to obtain a fair comparison between proposed and baseline systems.

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Load-Sensing Implements Model
The Load-Sensing system model is developed in Matlab/Simulink and it can be divided into six different submodels as shown in the top-level block diagram presented in Figure 3. Since the kinematic model has beenpresented and validated in previous publications [11] [12], its description will be left out of this paper, but thereader is encouraged to refer to mentioned references for more details. The operator model is a PI controller thatgenerates valve position commands based on a reference cycle profile. Additional inputs to the model include theexternal load (𝐹+,- ), which is estimated based on measured pressures, and the measured engine speed (𝑛* ). The
hydraulic components are modeled as follows:

Figure 3. Load-sensing model top-level block diagram

Pump and Main Line Models
The pump model is used to evaluate the effective flow and torque of the LS pump. The pump effective flow rate
is obtained with eq. (2):𝑄. = 𝑛* 𝑉. 𝛽. − 𝑄" ,𝑄" = 𝑓1𝑛* , 𝑝. , 𝛽. 3(2)where the 𝑉. is the pump maximum volumetric displacement, 𝛽. is the normalized fractional displacement and𝑄" accounts for the volumetric losses. The last term is obtained with an empirically derived loss model obtainedwith steady-state measurements and is a function of pump speed, pressure differential and displacement. Similarly,
the pump effective torque is obtained with:𝑀. =𝑉. 𝛽. 𝑝.
+ 𝑀" ,
2𝜋

𝑀" = 𝑓1𝑛* , 𝑝. , 𝛽. 3(3)where 𝑀" accounts for the hydromechanical losses. In both cases of eq. (2) and (3) the losses are estimatedaccording to the method proposed by Ivantysynova [13]. The pump fractional displacement can be obtained witha force balance in the pump control piston. If dynamic effects are neglected,

𝛽!/0 =

𝑥.

𝑥𝑝,𝑚𝑎𝑥=𝑝1" 𝐴. − 𝑝. 𝐴. + 𝐹".&'$(
𝐾".&'$( 𝑥/),(4)

where 𝐹".&'$( is the spring pre-load, 𝐴. is the control piston area, 𝐾".&'$( is the spring constant and 𝑥.,/), is thecontrol piston maximum stroke. The LS pressure (𝑝1" ) is the highest user pressure among the lines connected to5

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the pump. To reduce the number of parameters to be tuned, the above equation can be simplified to a proportionalpressure controller such that:𝛽!/0 =

𝑥.

𝑥𝑝,𝑚𝑎𝑥= 𝑘. ∙ (𝑝1" − 𝑝. + 𝑠)

(5)where 𝑠 is the so-called pump margin given by the spring pre-load. To avoid unrealistic dynamic behavior, afirst order transfer function is used to slow down transients. In this way
𝛽̇. =

1
1
𝛽!/0 + 𝛽.
𝜏.
𝜏.

(6)

where 𝜏. is the transfer function time constant. In addition, the pump displacement can be limited depending on
the simulated pump pressure, to mimic torque limiting controls usually implemented in these pumps to avoid
engine stall. The pump outlet pressure, is obtained with the pressure build-up equation accounting for all the flows
in and out of the line
𝑝̇. =𝐾
∙ 1𝑄. − 𝑄3,"4'$( − 𝑄3,566/ − 𝑄3,)&/ − 𝑄3,5#!7*- 3𝑉(7)where 𝑄3,"4'$( , 𝑄3,566/ , 𝑄3,)&/ , and 𝑄3,5#!7*- are the flows to each actuator out of the pump line.
Flow Control Valves and Compensators
The flow through work-port of the flow control valves is obtained considering the turbulent flow orifice equation.
Since the goal of this paper is to estimate power consumption, compensator dynamics are neglected. This can bedone because compensator dynamics are considerably faster than the actuator dynamics. In this way, the pressure
differential across the main spool is always equal to Δ𝑝3 = 𝑝. − 𝑝1" − 𝑠! where 𝑠! is the compensator spring
setting. In this way,
𝑄3 = 𝑘3 𝐴3 (𝑥3 )@Δ𝑝3(8)

where 𝑘3 is a constant flow coefficient, 𝐴3 (𝑥) is the opening area, which is a function of the spool position. To
reproduce the effect of the non-return check valve, the value of Δ𝑝3 is saturated to zero when it becomes negative.
Similarly, for the case of the return line
𝑄3,& = −𝑘3,& 𝐴3,& (𝑥3 )@𝑝)!- − 𝑝&*-#&$(9)

where the return pressure is assumed constant and equal to 4 bar and 𝑝)!- is the pressure in the actuator line
connected to the return line, according to the operator command.
Cylinders and Swing Motor
Similarly to the case of the pump outlet pressure, the pressure in actuator lines are also obtained with the pressurebuild-up equation. In the case of the linear actuators, the pressure in the bore chamber (𝑝8 ) and in the rod chamber(𝑝9 ) are evaluated as follows:
𝑝̇8 =
𝑝̇9 =

:
;!
:
;"

A𝑄8 − 𝐴8 𝑥̇ − 𝐾< ∙ (𝑝8 − 𝑝9 ) − 𝑄=,8 − 𝑄56,8 B,

𝑉8 = 𝐴8 ∙ 𝑥 + 𝑉0

(10)

A𝑄9 + 𝐴9 𝑥̇ + 𝐾< ∙ (𝑝8 − 𝑝9 ) − 𝑄=,9 − 𝑄56,9 B,

𝑉9 = 𝐴9 ∙ (𝑥/), − 𝑥) + 𝑉0

(11)

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where 𝑄8 and 𝑄9 can be either 𝑄3 or 𝑄3,& depending on the sign of the user command. 𝐾< is the internal leakage
coefficient, 𝑄=,8 , 𝑄=,9 are the relief valve flows (considered proportional to the chamber pressure when above therelief setting) and 𝑄56,8 , 𝑄56,9 are the laminar flows through the bleed orifice in the LS line. In this way, for thecase of the bore chamber:
𝑄56,8 = C

𝑘56 ∙ 𝑝8 ,0,

𝑝8 = 𝑝1"
𝑜𝑡ℎ𝑒𝑟𝑤𝑖𝑠𝑒(12)

Each actuator force can be obtained based on simulated chamber pressures and respective areas. In addition, a
Coulomb friction model is used to account for both static and dynamic friction.
The pressures in the swing motor line are modeled in a similar way, therefore:
𝑝̇8,"4 =𝐾
A𝑄 + 𝑄"4,8 − 𝑄=,8 − 𝑄56,8 B
𝑉8,"4 8(13)

𝑝̇9,"4 =𝐾
A𝑄 + 𝑄"4,9 − 𝑄=,9 − 𝑄56,9 B
𝑉9,"4 9(14)

where 𝑄"4,8 , 𝑄"4,9 are the swing motor effective flow rates to/from lines A and B, respectively. These effective
flow rates are calculated as follows:𝑛 ∙ 𝑉 − 𝑄"4,"𝑄"4,8 = C "4 "4𝑛>? ∙ 𝑉"4−𝑛"4 ∙ 𝑉"4>? ∙ 𝑉"4 − 𝑄"4,"𝑄"4,9 = C−𝑛, Δ𝑝"4 ≥ 0, 𝛥𝑝"4 < 0, Δ𝑝"4 ≥ 0, 𝛥𝑝"4 < 0

(15)

(16)

where 𝑄"4," = 𝑓(Δ𝑝"4 , 𝑛"4 ). Similarly, the swing motor effective torque, is given by:
𝑀"4 =𝑉"4 Δ𝑝"4− 𝑀"4," ,2𝜋

𝑀"4," = 𝑓(𝑛"4 , Δ𝑝"4 )(17)Reference Cycle and Model Validation
To verify the validity of the model of the hydraulic system, comparisons on measured kinematic quantities andpressures of the simulated machine are presented. A 90 degrees truck-loading cycle performed by an expert
operator has been selected. In this cycle, the machine digs into the ground, rotates 90 degrees while raising theboom and dumps the bucket content into a truck. The rationale for the selection of this cycle is twofold: first, it
has been widely used both in academia and industry for fuel consumption evaluation of excavators. Second, it
reproduces the most demanding operating conditions for the machine, in which four actuators are usedsimultaneously with maximum power demand.
Based on the available data, it is chosen to look at the actuators measured positions and pump outlet pressure. In
short, by comparing measured and simulated positions, it is guaranteed that the pump is delivering a similar flowrate in both experiments and simulation. Additionally, the validation of the pump outlet pressure ensures that theloads and throttling losses are well captured with the model. Therefore, it is inferred that the developed model isrepresentative of the LS pump power consumption.
Actuator positions are shown in Figure 4. Simulated results show a good agreement with measurements, therefore
demonstrating that the operator model is able to track the reference cycle well. The Load-sensing pump outletpressure validation is shown in Figure 5. A perfect match between measured and simulated LS pump outletpressure is somewhat challenging to obtain with the proposed model since even a small difference in operator
command can connect a different chamber to the LS line. Nevertheless, simulation results show a good agreementwith measurements.

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(a)

(b)

(c)

(d)

Figure 4. Simulated actuators positions with LS model. Boom (a), Bucket, (b), Arm (c), Swing (d)Figure 5. Load-sensing pump outlet pressure validation

PLANETARY GEAR CONFIGURATION SELECTION AND SYSTEM SIZINGEngine Downsizing
Although the proposed system allows any engine to be operated at the desired combination of torque/ and speed,
fuel efficiency can be further improved when smaller engines are used. This because a downsized engine will tend
to operate more often close to its peak torque, where fuel efficiency is higher and because the hydraulicaccumulator can be used to energy storage/release. Therefore, the starting point in the design process is to evaluate
the required engine size as well as the optimal engine torque/speed, such that the transmission can be designed tooperate the engine at the desired operating point.
To evaluate the required engine size, the developed model is used to evaluate the typical LS pump power demand
over a highly demanding cycle, as shown in Figure 6. From these results, it is clear the large variations in power
demand over time and therefore it is possible to size the engine to provide only the average power demand anduse the energy storage device to handle the heavy transients. For the cycle in analysis, the average demand is 22.8
kW, which is 37.5% lower than the original engine rated power of 36.5 kW. However, such a large engine size
reduction is not desired for three reasons: first, there is no guarantee that the cycle in analysis – even if it was anaggressive expert operator cycle – is the most demanding a machine will ever face in real life, meaning different
conditions and operators could require a larger average power. Second, this analysis considers only the LS power
demand and does not leave any extra power to accelerate the engine, for example should its speed drop during asudden load increase. Lastly, other functions in the machine like the tracks usually require a constant power
supply, and therefore it is not always possible to rely on the accumulator to achieve a higher power supply.

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Figure 6. LS Pump power Demand

The aforementioned reasons leave the room for the question on how to properly select the engine size. To answer
this, a proposed reasonable approach can be derived from the engine fuel consumption map. Diesel engines
typically operate with higher efficiency (or break-specific fuel consumption) near its peak torque, and not near its
peak power. Therefore, an engine can be selected to deliver the average power of the studied cycle at its peakefficiency point. This ensures optimal engine operation for a typical cycle, while also leaving some room tooperate at higher power demands should the operator require it. Nevertheless, this approach still requires ensuring
all the machine requirements are met, such as the power requirements of functions with constant demand like thetracks.
A typical Diesel engine fuel consumption map is shown in Figure 7. In this work, this map is scaled to thespecifications of the original engine in the baseline machine. The Figure shows that the peak engine power is at
2300 rpm, while the peak efficiency (and peak torque) is at around 1800 rpm. Therefore, assuming the smaller
engine will have similar characteristics, it is desired to select an engine capable of providing 22.8kW at 1800 rpm,
resulting in an engine with a maximum power of 29.2 kW, a 20% reduction when compared to the baselinemachine. Evidently, different engines will have different fuel consumption maps and the proposed approach may
not be as straightforward as presented here, requiring some iterations and analysis of different engines.
Nevertheless, it is reasonable to expect that other engines will share similar characteristics.
Besides operating the engine at its peak efficiency, the proposed approach also leaves the possibility of achieving
higher power demands. The selected engine can achieve up to 28% more power than what is seen on average in
the analyzed cycle, therefore ensuring that more demanding cycles can also be met, and that the accumulator canbe quickly charged. In addition, the reduction in power availability to the tracks is not as significant.

Figure 7. Typical Diesel Engine Fuel Consumption Map. Colormap is Break-specific Fuel Consumption (g/kWh)Planetary Gear Configuration Selection
With the engine size and preferred engine speed defined, it is desired to select the transmission components. A
challenge when designing any power-split transmission is to select the most appropriate configuration, meaning
to which one of the PGT shafts should engine, primary and secondary units be connected. Such selection is usually
done such that the primary unit speed is as close as possible to the full mechanical point, where the primary unit
speed is zero. Therefore, in this condition, most of the power is still be transmitted mechanically through the PGT.
The rationale for that approach is based on the fact that mechanical power transmission tends to be more efficientthan using a hydrostatic path, where the power flow is subject to total efficiencies of 2 hydrostatic units [14]. In

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the case of the proposed application, the job of placing the full mechanical point is simplified given the fact thatthe baseline machine runs with a constant engine speed. In this way, assuming no changes will be made to the LS
pump size, it is straightforward to select a configuration that will place the full mechanical point at an engine
speed of 1800 rpm and a pump speed of 2400 rpm. This can be done in an iterative process using equation (1) inwhich different gear ratios and configurations can be plugged in until the solution with the primary unit speed as
close to zero as possible is found for the given operating condition. It should be noted that different combinationsof configurations and gear ratios may lead to similar primary unit speeds. In that case, the PGT that requires
smaller hydrostatic units is preferred, which in the studied case led to the selection of the configuration shown in
Figure 2. In this case, the carrier is connected to the engine, the sun is connected to the primary unit and ring isconnected to the load. The sizing procedure for the hydrostatic units and accumulator is left out of this paper for
brevity, but the reader is invited to refer to [8] for details on the sizing method. However, it is important to highlightthat the system is sized such that the power of the original system can always be delivered to the LS pump, evenwhen the accumulator pressure is minimum. As a consequence, the proposed powertrain will have higher power
availability than the baseline machine. The final sizing is presented in Table 2. The table contains also other key
parameters of the system that will be detailed in the next sections.
Table 2. Sizing Results

Machine Parameter
Engine Power
Planetary Gear Standing ratio
Primary Unit Displacement
Secondary Unit Displacement
HP Accumulator Volume
HP accumulator pre-charge

Value
29.2 kW-2.5
12 𝑐𝑚@ /rev
21 𝑐𝑚@ /rev
6L
180 bar

ENGINE AND TRANSMISSION MODELSIn the case of the baseline system, the engine speed was measured and then used as an input to the model. On theother hand, in the case of the hybrid system it is desirable to develop a model that includes engine dynamics suchthat engine torque, speed and fuel consumption can be more accurately predicted. Applying a torque balance inthe engine shaft, the engine speed is obtained with𝜔̇ * =

1
1𝑀 − 𝑀!. + 𝑀!)&&'*& 3
𝐼* *

(18)

where 𝑀!)&&'*& is the torque at the carrier gear and 𝑀!. is the torque applied by the charge pump. The engine
torque is commanded by the engine governor, which in this case is represented by a PID controller trying to follow
a reference speed. To represent the dynamics associated with the turbocharger inertia and fuel injection delay, asecond-order transfer function is used to reproduce the results obtained by Tsai and Goyal [15]. In addition,
maximum engine torque is limited according to the engine wide-open-throttle curve and minimum torque islimited to 𝑀A&'! = 𝑓(𝑛* ) where the function was empirically derived by Heywood [16].
The transmission modeling followed the same approach used by Bertolin and Vacca in [8] but is reproduced here
for completeness. The nomenclature used in the modeling description is shown in Figure 8. The transmission
output shaft speed is evaluated with a torque balance, therefore𝜔̇ . = 𝜔̇ B =1
𝐼6#-

1𝑀&'$( − 𝑀B − 𝑀. 3

(19)where 𝑀&'$( is the torque at the ring gear, and 𝑀. is the load torque from the LS pump. Since 𝑛B = 𝑛&'$( , and
𝑛* = 𝑛!)&&'*& , equation (1) can be re-organized to evaluate 𝑛C , therefore defining all the speeds in the PGT.

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Figure 8. Transmission Model Variables

The torques at the carrier and ring gears are obtained based on the torque of the primary unit. Since the torquerelationship in the PGT is fixed,𝑀!)&&'*& = (𝑖% − 1)𝑀"#$ = (𝑖% − 1)𝑀C

(20)

𝑀&'$( = −𝑖% 𝑀"#$ = −𝑖% 𝑀C(21)However, the above equations do not consider any torque losses in the gearing. To account for those, an efficiency
map generated based on information provided by a partner company was used. The losses in the PGT are torqueand speed dependent. In this way, 𝑀!)&&'*& = 𝑓(𝑀C , 𝑛* , 𝑛&'$( ) and 𝑀&'$( = 𝑓(𝑀C , 𝑛* , 𝑛&'$( ).
The model of the hydrostatic path of the transmission evaluates the high-pressure accumulator pressure,accounting for its capacitance which is dependent on the current pressure. Therefore:1𝑝̇D. =(𝑄C + 𝑄B ).

E
𝑝%
𝑉
𝑉
Q 𝛾% S EFC
T . + 1'$*
𝐾 V
E
𝑝D.

(22)

where 𝛾 is the gas polytropic coefficient, 𝑝% is the accumulator pre-charge pressure and 𝑉% is the accumulator total
volume. In a similar way, the approach described by equation (22) is used to evaluate the pressure in the lowpressure side of the transmission (𝑝1. ) . The effective flow-rates of the primary and secondary units, 𝑄C and 𝑄B ,are obtained with a similar approach to the one described in equation (2). Similarly, the effective torques areevaluated with the same approach shown in (3).

Secondary Unit Displacement Control System
One of the challenges of the proposed architecture is the speed control of the output shaft. The common shaft
connecting secondary unit and LS pump has very low inertia when compared to the engine fly-wheel, but it canbe submitted to large and fast load variations as the LS pump displacement and outlet pressure rapidly changeover the cycle. Therefore, high-bandwidth is desired in the speed control of the secondary unit. However, suchbandwidth will be limited to how quickly the secondary displacement can be adjusted. To model these dynamics,
a method proposed by Grabbel and Ivantysynova [17] is used, which neglects swashplate inertia. In this way,𝛽Ḃ =1
𝑦 𝑘 @𝑝 .
𝑥.,/),,B 𝐴.,B "3 "3 D.

(23)

where 𝑥.,/),,B is the control piston stroke, 𝐴.,B is the control piston area, 𝑦"3 is the servo-valve position and 𝑘"3
is the servo-valve flow gain. The servo-valve dynamic response is modeled with a second-order model. In
addition, it should be noted that to increase the system response, the high-pressure line is used as power supply tothe control piston.

11

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SUPERVISORY POWER-MANAGEMENT CONTROLLER
The proposed architecture has the advantage of independent control of engine torque, speed and the LS pumpspeed. The purpose of the supervisory power-management controller is to control these states such that the stateof-charge of the accumulator is kept within the desired range, while ensuring that the LS pump can deliver therequired flow in the most efficient way possible.
The controller is designed such that the LS pump shaft speed is varied over time. The rationale for that is twofold:
first, it forces the pump to operate at higher displacements, making it is more efficient. Second, it allows anincrease in the speed of the primary unit, which results in a quicker charge of the accumulator during low power
demand moments of the cycle. To achieve that, it is proposed to use a closed-loop pressure control to determine
the desired output shaft speed. Therefore:𝑛B,&*A = 𝑘.,B (𝑝&*A − 𝑝. ), 𝑝&*A = 𝑝1" + 𝑠(24)where 𝑘.,B is a proportional gain. It is worth mentioning that the proposed system still considers a typical LS
pump with hydro-mechanical pressure feedback. In this scenario, the proposed approach has the following
advantages:••••

With the proper gain selection, the reference speed will always be close to the minimum required toensure that 𝑝. = 𝑝&*A , which then forces the LS pump to operate at higher displacements.
Since two independent controllers act on different control inputs, the proposed controller achieves afaster response when guaranteeing good tracking of the desired pump outlet pressure.
E-LS can be implemented without any changes to the LS pump, since the supervisory controller can
automatically reduce the pump margin by forcing the LS pump to flow saturate.
The only sensors required for implementation are LS pressure and pump outlet pressure.

To ensure that the accumulator is properly charged, the supervisory controller also changes the commanded power
to the engine. Using the derived simulation model, the following set of rules is proposed:
5 𝑘𝑊, 𝑝D. > 310 𝑏𝑎𝑟
𝑃*, !/0 = Y17 𝑘𝑊, 275 𝑏𝑎𝑟 ≤ 𝑝D. ≤ 310 𝑏𝑎𝑟
27 𝑘𝑊, 𝑝D. < 275 𝑏𝑎𝑟(25)

The idea behind this approach is to command an engine power above the average requirement for two reasons.
First, because the engine now also needs to provide power to account for the transmission losses. Second, because
it guarantees a quick recharge of the accumulator ensuring power availability. Once the pressure reaches a certain
level, then the power command is lowered to avoid constant high-pressure in the HP line, which can increase thevolumetric losses in the units. On the other hand, it should be said that this simplified approach is most-likely not
optimal, and improvements could be made.
Once the desired engine power is selected, an online optimization algorithm is used to find the optimal
combination of torque and speed that delivers the request. Based on the desired torque engine torque, the required
primary unit is torque is
𝑀C,!/0 =𝑀*,!/0 − 𝑀!.
(𝑖% − 1)

(26)

From (26), an inverse loss model is used to obtain the displacement 𝛽C,!/0 = 𝑓(𝑀C,!/0 , 𝑛C , 𝑝D. ) needed toachieve the required torque. To avoid sudden changes in the engine load, the displacement command signal to theprimary unit is filtered with a low-pass filter. In addition, the displacement of the secondary unit is limited in thecase the pressure in the HP line drops below the minimum threshold of 200 bar. A top-level block diagram of thecontrol approach is shown in Figure 9.

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Figure 9. Transmission System Control Top-level Block Diagram

IMPLEMENTATION WITH HYBRID SWING
Another benefit of hybridization is the storage of recovered energy, which is not possible with traditional LS
actuation. Based on the proposed architecture, a straightforward solution is to integrate the swing motor in thehydraulic path of the transmission, as presented in Figure 10. The swing is selected since it is the only actuatorthat can be switched to secondary control. The benefits of such approach are twofold: first, it allows the recoveryof kinetic energy from the swing system. Second, it eliminates the throttling losses associated with the swing
function in the LS line. Therefore, it is proposed to replace the fixed displacement swing motor with a variabledisplacement motor with secondary control. To quantify the benefits of the idea, the previously described model
was modified to include the swing motor in the hydrostatic path of the transmission.

SIMULATION RESULTS
Before evaluating the fuel consumption improvements of the proposed system, it appropriate to verify that it canactually deliver the same performance of the baseline machine. Ideally, this would be done experimentally by
evaluating the machine productivity in tons/liter, but in this simulation study this is done by evaluating how well
the proposed architectures follow the baseline cycle. The simulated outlet pump pressures are shown in Figure 11
from which is seen that for most of the cycle, the LS pump outlet pressure is the same. This shows that eventhough the pump speed is being varied over time, it does not change the ability of the LS pump to track the desiredload-pressure and therefore no significant changes in operator perception should be seen.
Boom

Arm

Bucket

Pe

Pmech
Phyd

PLS
Pacc
Psw

Primary Unit

Swing

P2

Secondary Unit

Figure 10. Proposed architecture with hybrid swing

13

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The discrepancies seen at 104s and 114s are most likely due to slight differences in the simulated operator
command, which may cause different chambers to connect to the LS line, for example if the command to valve
position becomes larger than the valve dead-band. Nevertheless, the results in Table 3 show that that all threesimulated systems present a nearly identical performance in terms of delivered energy.

Figure 11. Simulated LS pump outlet pressure

A comparison between the two proposed architectures is shown in Figure 12 and Figure 13. From Figure 12 it isseen that some swing kinetic energy is possible, although it is not that significant due to unit losses and friction.
Perhaps, more benefits would be seen in machines with larger inertia, but nevertheless it is clear from thesimulated pressures that recovery occurs, especially at 140s and at 152s when the pressure is clearly higher in thesystem with integrated swing, even though both systems were using the same power-management strategy.
Another important aspect to be highlighted in the transmission pressures is that both hybrid systems started andended the cycle with nearly the same pressure. Therefore, no free energy coming from an initial condition in theaccumulator pressure was used to accomplish the cycle and reduce fuel consumption estimation.

Figure 12. Simulated Transmission pressure differential and swing speed

The LS pump speed is shown in Figure 13. When the swing motor is in use, the pump speed is lower in the systemwith integrated swing (i.e., at 110s and 116s) therefore demonstrating the ability of the proposed controller toreduce properly slow-down the pump speed under low flow demand. It must also be pointed out that the speed
controller used is able to track the reference speed, despite low inertia and sudden load changes. Nevertheless,

14

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some spikes in speed are still seen (i.e., at 120s) when the LS pressure suddenly decreases, quickly reducing theload in the shaft.

(a)

(b)

Figure 13. Simulated Transmission Speeds for Power-Split Hybrid (a) and Power-Split + Swing Hybrid (b)A summary of the simulation results is shown in Table 3. Results show that the proposed transmission can reduce
the fuel consumption in up to 11.2%. This value can be extended up to 16.4% in the analyzed cycle when theswing is integrated in the hydrostatic path of the transmission, while respecting the same expert operator cycle.
The results can be supported by Figure 14 which shows the engine operating points. A significant change in engine
operating conditions allows the prime mover to always operate at peak efficiency. The average displacement of
the supply pump is also higher due to the implemented controllers, thereby increasing their efficiency.
Nevertheless, it should be noted that these gains are not that significant given the fact that the pump in the baseline
LS system also operated at a high average displacement. Lastly, the integration of the swing in the transmission
allows some energy recovery of the kinetic energy available, while also eliminating the associated throttlinglosses.
Table 3. Summary of Simulation Results

Parameter
Engine Downsizing
Average LS Pump Displacement
Delivered Positive Work
Fuel Consumption
Difference in Fuel Consumption

Baseline

PS Hybrid

PS + Swing Hybrid

-

-20%

-20 %

85.3%
388.6 kJ
91.27 g93.7%
377.0kJ
81.08 g90.2 %
388.2 kJ
76.34 g--11.2 %

-16.4%

CONCLUSIONSThis work introduced a novel hydraulic architecture by combining an efficient power-split transmission with stateof-the-art hydraulic actuation for high engine/hydraulics integration in an excavator. A simulation model has beendeveloped and validated for the baseline LS system of the reference machine, a 5 ton mini excavator. The model
was then used to design and evaluate the benefits of the proposed transmission in simulation.
Simulation results showed the feasibility of developing a design with high integration between engine andhydraulics for off-road vehicles. By allowing the engine to operate at its peak efficiency, up to 11.2% improvementin fuel consumption can be achieved, while still maintaining the reliability and user experience of state-of-the-art
hydraulics. The proposed architecture should also result in a lower up-front cost when compared to more radical
approaches. It should be noted that, even though this analysis considered a load-sensing architecture, similar
approach could also be followed with different architectures such as open-center systems.
In addition, the proposed system can be especially attractive to engine manufacturers, as it would allow an engine
design focused on a single point of operation. In this scenario, higher improvements can be expected.
Another advantage of the proposed system is the possibility of swing integration, which increases efficiency
improvements by reducing throttling losses and allowing energy recovery. When such approach is used,
improvements of 16.4% in fuel consumption are estimated for the reference cycle.

15

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(a)

(b)

(c)
Figure 14. Simulated Engine Operation of baseline system (a), Power-split Hybrid (b) and Power-split + Swing hybrid (c)REFERENCES
[1]
[2]
[3]
[4]
[5]
[6]
[7]
[8]

M. Vukovic, S. Sgro, and H. Murrenhoff, “Steam - A mobile hydraulic system with engine integration,”
ASME/BATH 2013 Symposium on Fluid Power and Motion Control, FPMC 2013, no. October 2013,
2013.
M. Vukovic, R. Leifeld, and H. Murrenhoff, “STEAM – a hydraulic hybrid architecture for excavators,”
Proceedings of the 10th International Fluid Power Conference, pp. 151–162, 2016.
K. Heybroek and M. Sahlman, “A hydraulic hybrid excavator based on multi-chamber cylinders andsecondary control – design and experimental validation,” International Journal of Fluid Power, vol. 19,
no. 2, pp. 1–15, 2018.
J. Zimmerman, E. Busquets, and M. Ivantysynova, “40% Fuel Savings By Displacement Control Leads
To Lower Working Temperatures - a Simulation Study and Measurements,” in Proceedings of the 52nd
National Conference on Fluid Power, 2011.
J. Zimmerman and M. Ivantysynova, “Hybrid Displacement Controlled Multi-Actuator Hydraulic
Systems,” in The Twelfth Scandinavian International Conference on Fluid Power, 2011, pp. 952–978.
R. Hippalgaonkar and M. Ivantysynova, “Optimal power management of hydraulic hybrid mobile
machines - Part I: Theoretical studies, modeling and simulation,” Journal of Dynamic Systems,
Measurement and Control, Transactions of the ASME, vol. 138, no. 5, 2016.
C. Williamson, “Power Management for Multi-Actuator Machines with Displacement Controlled
Hydraulic Actuators,” Purdue University, 2011.
M. Bertolin and A. Vacca, “An Energy Efficient Power-Split Hybrid Transmission System to Drive16

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The 11th JFPS International Symposium on Fluid Power 2020

[9]
[10]
[11]
[12]
[13]
[14]
[15]
[16]
[17]

Hydraulic Implements in Construction Machines,” ASME Journal of Dynamic Systems, Measurement
and Control, 2021.
R. Resch and K. T. Renius, “Continuously Variable Tractor Transmissions,” in Agricultural Equipment
Technology Conference, 2005, no. 29, pp. 1–37.
R. Kumar, M. Ivantysynova, and K. Williams, “Study of energetic characteristics in power split drives
for on highway trucks and wheel loaders,” SAE Technical Papers, vol. 116, no. May 2021, pp. 227–237,
2007, doi: 10.4271/2007-01-4193.
J. D. Zimmerman, M. Pelosi, C. A. Williamson, and M. Ivantysynova, “Energy Consumption of a LS
Excavator Hydraulic System,” in Proceedings of the International Mechanical Engineering PCongress
and Exposition, 2007, pp. 1–10.
E. Busquets and M. Ivantysynova, “Adaptive Robust Motion Control of an Excavator Hydraulic Hybrid
Swing Drive,” SAE International Journal of Commercial Vehicles, vol. 8, no. 2, pp. 568–582, 2015, doi:
10.4271/2015-01-2853.
M. Ivantysynova, “Energy Losses of Modern Displacement Machines - A New Approach of Modeling,”
in The seventh Scandinavian International Conference onFluid Power (SICFP’01), 2001.
B. Carl, M. Ivantysynova, and K. Williams, “Comparison of Operational Characteristics in Power Split
Continuously Variable Transmissions,” SAE Technical Paper Series, no. 2006-01–3468, 2006.
S. Tsai and M. R. Goyal, “Dynamic Turbocharged Diesel Engin Model for Control Analysis andDesig,” vol. 95, no. 1986, pp. 81–86, 2019.
J. B. Heywood, Internal Combustion Engine Fundamentals, 1st editio. MCGraw-Hill, 1988.
J. Grabbel and M. Ivantysynova, “An Investigation of Swash Plate Control Concepts for Displacement
Controlled Actuators,” International Journal of Fluid Power, vol. 6, no. 2, pp. 19–36, 2005.

NOMENCLATURE
𝐴8
𝐴9
𝐴.,B
𝐴.

Linear actuator effective area in the bore chamber
Linear actuator effective area in the rod chamber
Secondary unit control piston area
LS pump control piston area

𝐴3,&
𝐴3𝐹".&'$(

Effective flow area in the return port of the flow control valve
Effective flow area in the inlet port of the flow control valve
LS pump control piston spring pre-load

𝐼*
𝐼6#𝐾<
𝐾".&'$(Engine moment of Inertia
Output shaft moment of inertia
Cylinder internal leakage coefficient
LS pump control piston spring constant

𝑀C,!/0𝑀B𝑀!)&&'*&
𝑀!.
𝑀*,!/0𝑀*
𝑀A&'!

Desired primary unit torque
Secondary unit effective torque
Carrier gear torqueCharge pump torqueCommanded engine torqueEngine torqueEngine friction torque𝑀.

LS pump effective torque

𝑀&'$(
𝑀"
𝑀"#$
𝑀"4,"
𝑀"4

Ring gear torqueLS pump hydromechanical losses
Sun gear torque
Swing motor hydromechanical losses
Swing motor effective torque17

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𝑃*,!/0
𝑄C
𝑄B
𝑄8,"4
𝑄8
𝑄9,"4
𝑄9
𝑄=,8
𝑄=,9
𝑄H
𝑄56
𝑄.
𝑄"4,"
𝑄3,&
𝑄3
𝑉%
𝑉8," 𝑤
𝑉8
𝑉9,"4
𝑉9
𝑉0
𝑉1'$*
𝑉.
𝑉"4
𝑖%
𝑘.,B

Commanded engine power
Effective flow from/to primary unit
Effective flow from/to secondary unit
Swing motor effective flow rate in/out line A
Flow from/to the control valve from the A line
Swing motor effective flow rate in/out line B
Flow from/to the control valve from the B line
Flow through the relief valve connected to the bore chamber
Flow through the relief valve connected to the rod chamber
LS pump volumetric losses
Flow to the LS line
LS pump effective flow
Swing motor volumetric losses
Flow control valve return flow.
Flow from the pump to the users through the flow control valve
Accumulator total volume
Total volume in the swing line A
Total instantaneous volume in the linear actuator bore chamber
Total volume in the swing line B
Total instantaneous volume in the linear actuator rod chamber
Cylinder dead volume
Transmission high-pressure line volume
LS pump maximum volumetric displacement
Swing motor displacement
PGT standing gear ratio
Secondary unit LS pressure control proportional gain𝑘.

LS pump adjustment system effective proportional gain

𝑘"3
𝑘3,&
𝑘3
𝑛C
𝑛B,&*A

Servo-valve flow-gain
Flow control valves flow coefficient in the return port.
Flow control valves flow coefficient in the inlet port.
Primary unit speed
Secondary unit reference speed𝑛!)&&'*&𝑛*𝑛&'$(

Carrier gear speed in rpm
Engine speed
Ring gear speed in rpm

𝑛"#$
𝑛"4
𝑝%
𝑝D.

Sun gear speed in rpm
Swing motor rotational speed
Accumulator pre-charge
Pressure in the transmission high-pressure side𝑝8," 𝑤𝑝8
𝑝9,"4
𝑝9𝑝)!𝑝1"
𝑝.

Swing pressure in line A
Linear actuator bore chamber pressure
Swing pressure in line B
Linear actuator rod chamber pressure
Actuator chamber pressure
Pressure in the load-sensing line
LS pump outlet pressure

18

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𝑝&*-#&$𝑥.,/),,B

Pressure in the return line
Secondary unit control piston stroke𝑥.,/),

LS pump control piston maximum stroke

𝑥.
𝑥3
𝑦"3
𝛽C,!/0
𝛽B
𝛽!/0
𝛽.
𝜏.

LS pump control piston
Flow control valve normalized position
servo-valve normalized position
Desired primary unit displacement.
Secondary unit normalized displacement
LS pump fractional displacement for mechanical equilibrium at the control piston
LS pump normalized fractional displacement
LS pump displacement time constant

𝜔B
𝜔*
𝜔.

Secondary unit rotational speed in rad/s
Engine speed in rad/s
LS pump rotational speed in rad/s

Δ𝑝"4
Δ𝑝3
𝐾
𝑠
𝑠!
𝑉
𝛾

Pressure differential across the swing motor
Pressure differential across flow control valves
Fluid bulk Modulus
LS pump margin
Compensator spring pre-load
Volume at LS pump outlet
Ideal gas polytropic coefficient19

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Control of Air Bubble Content in Working Oil by Swirling Flow
Sayako SAKAMA*, Yutaka TANAKA**, Yasuhiro KODERA***, Yoshiaki KITAMURA*****

National Institute of Advanced Industrial Science and Technology
1-1-1 Umezono, Tsukuba-shi, Ibaraki 305-8560, Japan**
Department of Engineering and Design, Hosei University
2-33 Ichigaya-Tamachi, Shinjuku-ku, Tokyo 162-0843, Japan(E-mail: y_tanaka@hosei.ac.jp)***
Automotive Components Operations, KYB Corporation, Tokyo 105-6111, Japan****
Basic Technology R&D Center, KYB Corporation, Tokyo 105-6111, Japan
Abstract. Mixing of air bubbles in the hydraulic oil in hydraulic power transmission systems causes deterioration of
the dynamic characteristics of the system, erosion of component, accelerated deterioration of the hydraulic oil,reduction of heat dissipation, and vibration noise. In this paper, a method for control air bubble content by a swirling
flow of working oil has been proposed and developed. The amount of air bubbles in the oil flowing out of a bubble
separation device is adjusted by controlling the opening and closing of the valve on a vent port of the device. The air
bubble content from an output port of the bubble separation device is controlled by the opening time of the vent valve.
The validity of the proposed method was experimentally and numerically verified.
Keywords: Air bubble, Bubble separation device, CFD, Hydraulic oil, Swirling flow, Vent port

INTRODUCTION
A few percent of air are dissolved in hydraulic working oils under atmospheric pressure. When the pressure drops
locally in hydraulic power transmission circuits, the dissolved air appears in the working oil. Small bubbles due
to cavitation stagnate in the hydraulic working oil for a certain period of time, and it takes long time to dissolve
again at high pressure or to float and defoam. Air is also mixed from the surface of the hydraulic oil due to thesloshing phenomenon caused by the vibration of the hydraulic circuit and reservoir. Therefore, in the hydraulicworking oil in hydraulic power transmission circuits of automobiles, there are many opportunities for air to bemixed in, so 10% to 30% or more of air bubbles are mixed in the working oil. These air bubbles in the hydraulicworking oil cause many troubles in hydraulic equipment and are called the third contamination that follows thesolid fine particles and water in the oil [1].
Air bubbles in the hydraulic oil reduce the rigidity of the oil and the system response due to sponge effect [2]. It
also causes vibration noise [3] and equipment damage [1]. In our previous studies [4], the amount of dissolved air
in hydraulic oil can be reduced by actively removing the air bubbles in the oil. As a result, it has been confirmedthat the occurrence of cavitation can be reduced. Since one of the causes of failure of hydraulic elements is erosion
due to cavitation, reduction of the air bubble content in hydraulic oil prevents damage to the equipment. However,to reduce the performance of the system, the quantitative index of how much air bubbles are contained in thehydraulic oil and how long it takes to reach a failure has not been experimentally clarified. The reason is that it isextremely difficult to accurately measure and control the air bubble content in the hydraulic working oil in realtime. To measure the amount of air bubbles in oil, mechanical, optical or electrical methods are available. Schrank
et al. [5] compared several methods to determine the amount of mixed air in oil in terms of difficulty of application,accuracy and influence of the measuring systems on the flow in the circuit. Sakama et al. [6] was proposed andevaluated the method for calculating the air volume fraction from measurements of the fluid’s effective bulk
modulus.
Conventional techniques for removing air bubbles in the hydraulic oil have been passive methods such as releasing
air from the surface of oil in the reservoir and adding anti-foaming agents to the hydraulic oil. However, in recent
automobiles, the frequency of intermittent operation including stopping the engine is increasing to improve fuelefficiency. The hydraulic pump connected to the engine is also required to start and stop frequently, and the air
bubbles in the hydraulic oil stagnant at various parts of the hydraulic circuit and gather to form a large air mass.
These air bubbles cause noise at the beginning of startup and instability of the operation of the hydraulic system.
Quantitative elucidation of the effect of air bubble content in the hydraulic oil on the hydraulic component is anurgent issue.
In this paper, a method for control air bubble content by a swirling flow of working oil has been proposed andbubble content adjustment system has been developed. A bubble separation and elimination device from theworking oil [7][8] has been used to control the air bubble content. The amount of air bubbles in the oil flowing

1

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out of the bubble separation device is adjusted by controlling the opening and closing of the valve on a vent portof the device. The air bubble content from an output port of the bubble separation device is controlled by the
opening time of the vent valve. The validity of the proposed method is experimentally and numerically verified.

BUBBLE SEPARATION DEVICE BY SWIRLING FLOW
Structure and principle of the bubble separation device is shown in Figure 1. Working oils with entrained air
bubbles tangentially flow into a tapered tube chamber from two inlet ports and form swirling flows that circulate
the fluid through the flow passage. The swirling flow accelerates to the downstream. Small bubbles are trappedand collected near the vicinity of the central axis in the tapered tube in a short time of about 100 ms by the swirling
flow [9]. When some back pressure is applied on the downstream side of the device, the collected bubbles arepushed out from the opposite side through a vent port. The amount of air bubbles to be separated and removed
from the working oil can be adjusted by providing a valve for opening and closing at downstream side of the ventport and controlling the opening and closing time of the valve. By using this principle, it is possible to constructa system that controls and adjusts the bubble content in the hydraulic oil on the downstream side of the device.

Figure 1. Principle and structure of bubble separation device

DESIGN AND MANUFACTURE OF BUBBLE SEPARATION DEVICE
The goal of the method verified in this study is to adjust the air bubble content in the oil in the range of 1 to 40%.
The performance of the bubble separation device must be sufficiently ensured within this bubble content range.
The performance of the device depends on the shape and dimensions of the inlet ports and the diameter of ventport [10]. In the design of the bubble separation device, the selection of shape parameters under the flow conditions
is very important to realize the method of adjusting and controlling the bubble content in the working oil.
The numerical simulation for the swirl flow in the bubble separation device was carried out using CFD software
STAR-CCM+. Geometry of the bubble separation device is shown in Figure 2. To select the optimum shapeparameters of the bubble separation device, the numerical analysis was performed under the conditions of air
particle diameter of 0.1 mm, bubble content in the hydraulic oil of 20%, 30% and 40%. A density of the oil is 0.85
g/cm3, a kinematic viscosity is 9.19 mm2/s, and a flow rate is 30 L/min.

Figure 2. Geometry of bubble elimination device

2

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Figure 3 shows typical simulation results of air particle content in the bubble separation device for bubble content
of 20%, 30%, and 40% respectively. It was confirmed that the air particles ware separated and removed from thevent port side of the bubble separation device under all conditions. As a result of numerical analysis, the optimum
shape parameters of the bubble separation device ware selected as tabulated in Table 1.

Figure 3. Numerical simulation of air particle content in bubble separation device
Table 1. Shape parameters of the bubble separation device

Shape parameters

Dimensions [mm]Inner diameter of tapered tube D1

30.0

Inner diameter of downstream D2

16.0

Inner diameter of vent port D3

4.4

Total length L

150.0

Inlet tube length L1

20.0

Tapered tube length L2

30.0

Downstream tube length L3

100.0

Width of inlet port w

2.18

Height of inlet port h

4.35

Figure 4 (a) shows the structure and dimensions of the bubble separation device designed and manufactured forthe experiment, and Figure 4 (b) shows a photo of the transparency device manufactured. The experimental bubble
separation device was designed separately so that the dimensions of the inlet ports and the diameters of the outlet(a) Dimensions of experimental bubble separation device(b) Transparency type of bubble separation device

Figure 4. Bubble separation device designed and manufactured for the experiment3

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tube and the vent port could be adjusted according to the experimental conditions. Each element designed by
dividing was inserted into a transparent housing, fixed from both sides, and integrated. To visualize the flow insidethe device, the housing and some of the elements were made of transparent acrylic.

PRELIMINARY EXPERIMENT
Preliminary experiments were carried out to confirm and verify the performance of the bubble separation and the
bubble content control method of the manufactured device.
Figure 5 shows a configuration of hydraulic circuit used in the preliminary experiment. The bubble separation
device was installed in the oil circulation circuit, and a manual needle valve was installed on the vent port side ofthe device and a Coriolis flow meter was installed on the outlet side of the device. Air bubbles were forcibly blown
into the working hydraulic oil from the delivery side of the pump by an air compressor. The pressure at the inlet
line was set at 0.5 MPa. The needle valve was manually adjusted while mixing air into the oil. The density of theoil on the downstream side of the bubble separation device was measured with respect to the opening rate of theneedle valve and the air content rate in the oil was estimated by the Coriolis flow meter. The volumetric flow rate
was 12.4 L/min in case of closed the needle valve. The oil temperature was kept at 40 1 ℃.
Figure 6 shows the results of the preliminary experiment. Due to the specifications of the needle valve, it isdifficult to precisely adjust the flow rate to the vent port depending on the valve opening. The bubble removal rateof the bubble separation device increased by slightly opening the needle valve. As a result, the air bubble content
on the outlet line of the device sharply decreased with respect to the opening rate of the needle valve. However,Bubble separation device
Vent line
Needle valve

Outlet line
Coriolis flow meter

Manual control

Inlet line

Air compressor

M

Figure 5. Experimental hydraulic circuit for manual control of air bubble content

Figure 6. Air bubble content at outlet line by manual control of the needle valve at the vent line

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The 11th JFPS International Symposium on Fluid Power 2020the validity of the basic method of controlling the amount of air bubbles on the outlet side was confirmed byadjusting the opening of the valve on the vent port side of the device.

BUBBLE CONTENT ADJUSTMENT EXPERIMENT
Based on the results of the preliminary experiment, an on-off solenoid valve was installed on the vent port andexperiments were carried out to adjust the air bubble content in the oil on the outlet line by controlling the valve
switching time and timing.
Figure 7 shows the configuration of the hydraulic circuit used in the adjustment experiment. A bubble separation
device was installed in the same circulation circuit as in Figure 5, and the Coriolis flow meter was installed on theinlet line and an in-line impedance type of the bubble content measurement sensor was installed on the outlet line.
The bubble content of each of the inlet and outlet lines of the device was measured in real time and stored in adata logger. Air was mixed into the oil by opening the suction side of the pump to the atmosphere. The on-off
solenoid valve was installed on the vent line of the bubble separation device. The opening and closing time of theon-off solenoid valve was controlled by a sequencer using a relay according to the deviation between the output
value of the bubble content sensor on the outlet line and the target value of the bubble content.

Data logger

Bubble separation device
Vent line
On/Off valve

Outlet line
Inline aeration meter

Relay

Inlet line
Sequencer

Coriolis flow meter
M

Target air content

Figure 7. Experimental hydraulic circuit for automatic control of air bubble content

Figure 8 shows a typical example of the adjustment results of the air bubble content. The measured pressure andvolumetric flow rate at the inlet line were 0.4 MPa and 11 L/min, respectively. The air bubble content of the inlet
line was set to about 9%, the air bubble content of the outlet line when the vent line was closed was set to about
28%, and the target value of the bubble content was set to 18%. The oil temperature was kept at 37 ℃. The cycleof opening and closing time of the on-off solenoid valve was adjusted to 1 s, and the duty ratio was adjusted in
the range of 10% to 100%. From the measurement results, if the duty ratio was small, the adjustment accuracy
and response slightly decreased. On the other hand, the average value of the adjustment results from 19.2 s to 25
s after settling near the target value is 16.8% to 19.7% at each duty ratio. This indicates that the relative error withrespect to the target value is within ± 9.4%.

CONCLUSIONSIn this paper, we proposed and developed a bubble amount control system that can arbitrarily adjust the amountof bubble contained in hydraulic oil using the technique of separating air bubbles in oil by swirling flow. We also
experimentally verified the validity of the control method. It was experimentally confirmed that the air bubble
content on the outlet line of the device can be controlled by adjusting the flow rate on the vent line of the bubble
separation device. In these experiments, it was also confirmed that precise adjustment according to the duty ratio
was difficult because the response of the on-off solenoid valve on the vent line was slower than the response of
the separation performance of the bubble separation device. In the future, it will be necessary to consider
countermeasures using high-speed proportional valves instead of the on-off solenoid valve.

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Duty

↓ Start29

Air bubble content [%]Average

Duty比=10%10%19.7%
Duty比=20%20%
18.8%
Duty比=30%
30%
17.9%

27

Duty比=40%40%
16.8%
Duty比=50%
50%
17.3%

25

Duty比=60%
60%
18.0%

23

Duty比=80%
80%
18.6%

Duty比=70%70%18.6%
Duty比=90%90%18.6%
Duty比=100%
100% 18.6%

21

Relative error
9.4%
4.4%
0.5%
6.7%
3.9%
0.0%
3.3%
3.3%
3.3%
3.3%

19
Target17
Average interval

15
0

5

10

15

20

25

Time [s]Figure 8. Adjustment results of air bubble content at outlet line by automatic control of vent line

ACKNOWLEDGMENTS
We express appreciation the technical support, Mr. Iori Masuhara of the master course student in the Graduate
School of Hosei University. This research was supported by the JSPS KAKENHI Grant Number 26420086 and
18K03916, the Precise Measurement Technology Promotion Foundation the MIKIYA Science and Technology
Foundation, and a grant-in-aid from the Ministry of Economy, Trade and Industry, Japan. The numericalsimulation was carried out by using the facility and software at the Research Center for Computing andMultimedia Studies, Hosei University.

REFERENCES
1. CANTER, N., Foam, Dealing with a Persistent Problem, Tribology & Lubrication Technology, December 2015, pp.24-38.
2. GHOLIZADEH, H., BURTON, R., SCHNOENAU, G., Fluid Bulk Modulus: A Literature Survey, International Journalof Fluid Power, 2011, Vol.12, No.3, pp.5-15.
3. ERICSON, L., PALMBERG, J., Measurement of Free Air in the Oil Close to a Hydraulic Pump, Proc. 7th International
Symposium on Fluid Power TOYAMA, 2008, P2-12.
4. GOTO, H., SAKAMA, S., SUZUKI, R., TANAKA, Y., Reduction of Cavitation Damage by Elimination of Bubbles in
Oil Reservoir, Proc. FLUCOME2013, 2013, OS1-02-4.
5. SCHRANK, K., MURRENHOFF, H., STAMMEN, C., Investigation of Different Methods to Measure the Entrained Air
Content in Hydraulic Oils, ASME/BATH 2014 Symposium on Fluid Power and Motion Control, Bath, United Kingdom,2014.
6. SAKAMA, S., KITAZAWA, Y., SUGAWARA, Y., TANAKA, Y., Estimating the Air Volume Fraction in Hydraulic Oil
by Measuring the Effective Bulk Modulus, 11th IFK International Conference Proceedings, 2018, Vol.3, pp.79-85.
7. SAKAMA, S., GOTO, H., HIGASHI, H., SUZUKI, R., TANAKA, Y., Air Bubble Separation and Elimination fromWorking Fluids for Performance Improvement of Hydraulic Systems, Proc. IFPE 2014, 2014, Paper No. 27.1.
8. TANAKA, Y., KISHI, Y., SAKAMA, S., Selective Separation of Air Bubbles from Working Oil by Bubble Elimination
Device, Proceeding of the 10th International Conference on Fluid Power Transmission and Control (ICFP2021), Hangzhou
in China, 2021, S2401, pp.703-706.
9. TANAKA, Y., SUZUKI, R., ARAI, K., IWAMOTO, K., KAWAZURA, K., Visualization of Flow Fields in a Bubble
Eliminator, Journal of Visualization, 2001, Vol.4, No.1, pp.81-90.
10. SAKAMA, S., TANAKA, Y., SUZUKI, R., Optimization of Bubble Eliminator by Numerical and Experimental
Investigation, International Journal of Automation Technology, 2012, Vol.6, No.4, pp. 418-425.

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Pulse Tests on Additive Manufactured Valve Blocks –Damage Analysis and New Design Possibilities
Sebastian DEUSTER*, Stefan AENGENHEISTER*, Gunnar MATTHIESEN*, Katharina
SCHMITZ**Institute for Fluid Power Drives and Systems
RWTH Aachen University
Campus-Boulevard 30, Aachen, 52074 Germany(E-Mail: sebastian.deuster@ifas.rwth-aachen.de)

Abstract. Additive manufacturing processes are currently experiencing an enormous upswing in different areas. The
demand for more efficient and powerful hydraulic systems has risen rapidly in recent years. Due to cost savings andresource conservation, there are always efforts to reduce the power-to-weight ratio of hydraulic components.
Especially in areas where weight savings lead to large performance savings, such as the aerospace industry,
additively manufactured components in hydraulic systems are becoming increasingly popular. This paper deals withthe testing of additively manufactured hydraulic manifold blocks with regard to their strength values. For thispurpose, the components are experimentally loaded with pressure pulses until they fail. Subsequently, the cause of
fracture is to be determined by means of a surface examination and thus a statement made about the suitability ofadditively manufactured components in hydraulic systems. These considerations show the potential of additive
manufacturing for weight and material savings in hydraulic applications.
Keywords: additive manufacturing, pulse tests, material testing, design optimization, fatigue strength

INTRODUCTION
Due to their potential in lightweight construction and the possible component complexity, additive
manufacturing processes are currently experiencing a tremendous upswing in all possible areas. The demand forincreasingly efficient and powerful hydraulic systems has risen rapidly in recent years. As a result of costsavings and resource conservation efforts are constantly being made to reduce the power-to-weight ratio of
hydraulic components. Especially in areas where weight savings lead to large power savings, such as theaerospace industry, additively manufactured components are experiencing a growing popularity in hydraulicsystems. In addition to the material savings which can be achieved, the advantages of additive manufacturing
also lie in the freedom of design. In its current state, additive manufacturing is an established tool especially for
rapid prototyping. Nonetheless, series production is realized in some applications which are exposed tooperational stress such as valve blocks in aerospace. The most common method for manufacturing additive
components is selective laser melting (SLM), which is based on using a laser to melt layers of metal powder.
In hydraulic line systems, for example, pressure loss savings can be achieved by adapting the geometry to thefluid flow. However, material parameters such as fatigue strength of these products are difficult to record due totheir selective design and are therefore often not sufficiently known.
The task in this paper is to test and compare additively and conventionally manufactured hydraulic distributor
blocks with respect to their strength values. For this purpose, the components are experimentally loaded withpressure pulses until their failure. Afterwards the cause of the fracture is determined by means of a surface
investigation and thus a statement is to be made about the suitability of additively manufactured components inhydraulic systems. These considerations demonstrate the potential of additive manufacturing for weight andmaterial savings in hydraulic applications. It is also possible to investigate whether the strength of the additively
manufactured blocks can be increased by minor geometry optimizations.
All blocks are tested simultaneously in a developed test rig. The additive manufactured blocks are produced by
selective laser melting (SLM). A machining of the threads and holes of these components is necessary toachieve the same function and functional surfaces as the components of the conventionally manufactured
blocks. To cover statistical uncertainty three versions of the manifold block, nine units are tested. Initially, theblocks are loaded with static pressure. The pressure is increased by 10 bar at fixed intervals until a maximum
pressure of 700 bar is reached. The pulse loading of the specimens follows the static tests. As test methods forthe pressure pulse, standards DIN EN 2624 [10] and DIN EN ISO 19879 [12] are applied. The pressure level ofthe pulses is above the range of conventional hydraulics between 500 and 600 bar. The test is stopped at block
failure, caused by visible damage or leakage due to small cracks in the component. The damage is detected bymeans of a visual inspection.

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Another damage analysis is carried out after the dynamic tests. Material tests are conducted to evaluate thematerial structure of the components. Non-destructive and destructive methods are used. In addition, thegeometry changes resulting from the pulse loads are examined. The goal is to analyze the differences between
the additively and conventionally manufactured valve blocks with regard to material failure. This allows studies
to be carried out in subsequent investigations on process optimization in the manufacturing of additive
manufactured hydraulic components, which deal with the design and structure.

STATE OF THE ART
Currently, conventionally manufactured components are mostly made from wrought material using cutting
methods. This type of process induces a high loss of non-reusable material. In this aspect, additive processes
have the advantage of involving minimal post-processing that could lead to loss of material. In accordance with
DIN 8580 [1], additive manufacturing is considered a primary shaping process. Additive manufacturing isestablished as a reliable method for rapid prototyping.
Another advantage of additive manufacturing is the use of high-temperature resistant materials such as
Inconel 718 [2]. This material is very difficult to form and machine so that shaping by conventional
manufacturing methods is only possible to a limited extent. In addition, a functional integration of components
can be carried out. This can reduce the assembly complexity and the number of joints in an assembly.
For loaded components however, additive manufacturing has not been applied until recently due to the lack ofcorresponding data and experience [3]. In hydraulic applications, the potential exists above all to achieve a
reduction in weight and optimizations of flow with the aid of components manufactured using additive
manufacturing. To this end, suitable components in applications are constructively optimized with regard to thepossibilities of additive manufacturing in order to exploit the advantages. Notable applications in recent years
include a valve block in the Airbus A380 and the series production of additively manufactured components at
Liebherr-Aerospace [4],[5]. Another example is an additive manufactured cartridge valve from Parker Hannifin.
An additive geometry optimization in this example led to a weight reduction of 42% and at the same time theflow rate was increased by 35% while maintaining the pressure drop of the valve. [6]Among various methods of additive manufacturing, selective laser melting (SLM) is most commonly used.
Therefore a laser is used to melt layers of metal powder. After one layer is melted and thereby bonded, more
powder is added, and the melting process is repeated for the next layer. At the end of the process, excess powder
can be removed and reused [3], [7]. Possible materials for the SLM process include titanium, aluminum,magnesium, and stainless steels [7]. In comparison to injection molding, SLM is advantageous because it avoids
the costly manufacturing of a mold. Further notable benefits are the automated manufacturing on the basis of a
CAD file and various options for reducing material usage. Some possible options to reduce material usage are
the afore-mentioned removal of excess material, lightweight design for the interior structure, and taking
advantage of the design flexibility to reduce material usage in specific locations of the component [8].
Up to this point, mechanical experiments on additively manufactured components have been limited to tensile
and notch impact testing. Predominantly, titanium alloys as well as steel-, nickel-, and aluminum-based alloys
were tested. As evident from Figure 1, the manufacturing method has a significant influence on the dynamic
durability of the component. The component quality is dependent on layer thickness, laser power, and theselected post-processing method. Furthermore, it has been shown that pre-heating the powder can reduce
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Laser: HIP at 920°C/2 h/100 MPa, processed surface
Laser: no treatment, unprocessed surface
Laser: 3h at 650°C, unprocessed surface , R = 0,1
Laser: no treatment, processed surface
Laser: 4h at 650°C, processed surface
MMPDS 2010: 3 inch cast, processed surface, R = 0,1
MMPDS 2010: forged, annealed, processed surface, R = 0,1
MMPDS 2010: forged, precipitation hardened, processed surface, R = 0,1
EBM: no treatment, processed surface
EBM: no treatment, unprocessed surface
EBM: HIP at 920°C/ 2h, processed surface
Laser metal forming, laser wire feed: heat treated and unprocessed surface
EBM: optimized, no treatment, unprocessed surface, R = 0,1
EBM: optimized, processed surface, R = 0,1
EBM: no treatment, processed surface, R = 0,1
Figure 1. Tensile strength dependency on manufacturing method [9]Standardization of Pulse Testing
Concerning pulse testing, several standardizations exist. The following three standardizations will be outlined
more closely: DIN EN 2624 [10], DIN ISO 6772 [11], and DIN EN ISO 19879 [12].
Experiments based on DIN EN 2624 are designated for the testing of passive components. A pressure increase
rate is defined between two points (10% and 90% of maximum pressure respectively) that may not exceed
14,000 bar/s. One test consists of 300,000 cycles at a rate of one to five cycles per second. A constant fluid
temperature equivalent to ambient temperature is maintained throughout the experiment. Testing is concluded
by inspecting the component for fissures and structural failures.
DIN ISO 6772 is a standardization for the testing of tubes, pipes, and screwed components. DIN ISO 6772
differs from DIN EN 2624 in regard to pressure profile and fluid temperature. One pressure maximum above
110% of the operating pressure is permitted within the first 15% (timewise) of each cycle. The maximum
pressure in relation to the operating pressure and depends on the pressure class and the nominal outside diameterand is defined within the standardization. The pressure increase rate is defined in the same way as in
DIN EN 2624 and its upper and lower limits are likewise defined within DIN ISO 6772. Before any test, it isnecessary to keep the system at testing temperature for one hour. Throughout the testing cycle, the temperature
is varied in a defined sequence.

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DIN EN ISO 19879 serves as a standardization for the testing of metallic pipe connections, screw-in parts,and flanges. The basic requirement is for the tested component to be uncoated and have a hardness between
35 HRC and 45 HRC. The maximum pressure may not exceed 133% of the operating pressure and the pulse
frequency must have a value between 0.5 Hz and 1.25 Hz.

APPROACH AND TESTING
For the investigations on the test bench, three different variations of the aluminum valve blocks were examined.
Figure 2 shows the different types. A conventionally manufactured block (1), an additively manufactured replicaof the conventionally block (2) and an additively manufactured version with reduced material usage (3) were
subject of the testing.
1

2

3

Figure 2. Tested valve block variants

Block 1 was made from 6061-T651 aluminum at SUN Hydraulics [13]. Blocks 2 and 3 were additively
manufactured at Aachen University of Applied Sciences (FH Aachen) using an AlSi10Mg alloy with a layer
thickness of 50 µm. The parts were produced by a selective laser melting process.
Due to the limited amount of available valve blocks, an initial static test was conducted. Using this test, apreliminary estimation of the upper pressure limit for the subsequent dynamic testing could be obtained.
In Figure 3 the test rig is shown. A radial piston pump (1) provides the necessary static pressure. The suction
line is prefilled by a gear pump. Pipes guide the oil to the test specimen carrier where the fluid flow is
distributed between up to six test blocks (3). The correct pressure and a uniform pressure load distribution are
ensured by the use of three pressure sensors and two pressure limitation valves within the valve block.
The pump (1) provides a steady fluid flow. In order to maintain a constant pressure, it is necessary to lead theaccumulating oil back to the tank in a controlled manner. Pressure oscillations are damped by the accumulator(4).
The static pressure has an initial value of 70 bar that is maintained for ten minutes. After each test pressure thatis conducted without any mechanical failure, the pressure is increased by 10 bar. This procedure is carried out
until mechanical failure of the block or as soon as the defined upper limit of 700 bar is reached. Pressures above
700 bar would put the remaining components of the test rig at risk and must therefore be avoided.
The locations of the sensors are indicated in Figure 3. A 3/2 directional valve (2) controls the pressure pulses byconnecting the tested blocks to the tank and accumulator (4) in alternation. Pressure sensors (Z1, Z2) on thebottom of the pulse block are used to monitor the intensity of the pressure pulses. Pressure sensor S determines
the pressure in the accumulator and defines the connection between pump and test blocks.
The returning line from the directional valve to the tank includes a filter (6) that removes contaminations. The
pulse block contains a pressure relief valve (5) to ensure that the pressure does not exceed the maximum
permitted value. The pressure at the back end of the test specimen is measured in Z2 and continuously comparedwith the values from Z1. This ensures that all blocks are exposed to the same pressure. The flow directions aswell as the motor speed can be controlled. Most of the components are connected by flexible hoses, whose
deformation enables the compensation of the pressure variations.

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U

P Z2

Test
Blocks
P Z1

U

PS

U

co n tro ller

M

Figure 3. Test bench setup

pmax
90% pmax

135
120
115
110
100
75
50

10% pmax
2%

Percentage of nominal pressure

Among the afore-mentioned standardizations, DIN EN 2624 [10] was selected. This standard was used because
the blocks to be tested correspond to the passive hydraulic components described in the standard. Components
such as liquid distributors are mentioned here. Furthermore the pressure curve can be represented by the test
bench. The pulse frequency can also be covered by the test bench.
The corresponding pressure profile is given in Figure 4. The framework conditions of the standard have alreadybeen described in the section ‘standardization of pulse testing’.

25
0

15

5

50

50

100
Percentage of load cycle

Figure 4. Pressure profile DIN EN 2624 [10]Because the pulse frequency is not well defined within the standardization, a frequency of 1 Hz was chosen sothat all three standardizations previously described are satisfied. An HLP ISO VG 46 was used as thepressurized fluid [14]. To calibrate the test bench and remove air from the system, approximately 100 pulses at
50% of the testing pressure were applied before the experiment. The initial pressure of the dynamic test was
defined in accordance with the results of the static test. To achieve control of the system, the sensor
measurements are evaluated with the measurement computer. The sensors S and Z2 provide the current pressurein pulse block, accumulator, and distributor block. As soon as sensor S indicates that the testing pressure has
been reached in the accumulator, the accumulator is connected to the tested valve blocks for 50% of one impulse
duration. At the end of this time interval, pulse block and valve block are separated again. Pressure builds up in
the accumulator and the cycle is repeated. The testing pressure can be adjusted by changing the pressure atwhich the pulse block switches the flow direction. The frequency is a function of the pressure build-up time andincreases as the motor speed increases. [10]Figure 5 illustrates the pressure course during the pulse test. Recorded are the pressure of the end of the test
specimen carrier (Z2) and the accumulator pressure (S). In the upper diagram respectful 20 seconds of the pulse

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testing are shown. The pulse pressure was 550 bar. It can be seen that a uniform and constant course of the pulse
length and frequency was achieved. The average pulse frequency of this recording is 0.95 Hz.
The figure below shows a detailed representation of the pressure curve over a period of one second. It can beseen how the pressure in the accumulator drops abruptly when the directional control valve is opened and theaccumulator is connected with the test blocks. Immediately afterwards a pressure peak (Z2) of 564 bar in the test
specimen carrier occurs. The oil volume of the accumulator fills the test specimen carrier and the test blocks,
before the valve opens in the direction of the tank. The pulse length is about 0.5 seconds. The pressure build uptime is 28,305 bar/s and is therefore within the validity range of DIN EN 2624 [10].

Figure 5. Pressure distribution and close up

INTERPRETATION OF TEST RESULTS
The purpose of the static tests was to estimate the stress limit of additive manufactured in comparison toconventionally manufactured valve blocks. All three test specimens were able to withstand the upper limit of
700 bar under static pressure. From this test, it was concluded that the initial dynamic pressure could be
increased to a larger value than the preliminary limit of 210 bar. [13]For the dynamic tests, it is essential to consider the geometrical boundary conditions. The threads on the upper
part of the valve block are incomplete as they are interrupted by the fluid duct. To ensure the sealing of thethreads, a stopper must be inserted. The stopper is at risk of unfastening under dynamic stress, so a threadlocking fluid was employed to increase the time until the sealing was breached.
During the test cycle, it proved to be impossible to maintain the desired frequency of 1 Hz. The volume flow of
the pump could not refill the accumulator within a short enough time, so it took longer to regain the testingpressure after each impulse. Other factors such as oil temperature, amount of simultaneously tested valve
blocks, and residue air in the system can complicate the matter. The complication was solved by manuallycontrolling the pump during the test cycle. The frequent changing of the valve blocks caused more air to enter
the system and further complicated a reliable pressure control. Ultimately, the valve blocks were loaded withfrequencies between 0.7 Hz and 1.1 Hz. As for the peak pressure, a sufficient accuracy was achieved. For atarget pressure of 550 bar, the actual pressure had a value between 548.1 bar and 553.1 bar.
The static tests showed that an initial pressure far greater than the stress limit is necessary to cause a failure.
Therefor for the first pulse tests, a pressure of 600 bar was selected. The additively manufactured valve blockwith optimized geometry lasted for less than 5% of the impulses of the additively manufactured valve block
without optimization and was therefore excluded from test series 2 and 3. Test series 2 and 3 were carried out at
550 bar and 500 bar respectively.
The conventionally manufactured valve block exhibited the expected behavior; as the pressure was decreased,
the amount of impulses endured increased. For the additively manufactured valve block however, the mount of

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The 11th JFPS International Symposium on Fluid Power 2020endured was roughly the same for all tested pressures. The results were summarized in a Wöhler diagram in
Figure 6.
625

Pressure in bar

600
575
550
525
Additively manufactured
Conventionally manufactured
Logarithmic trendline (additively)
Logarithmic trendline (conventionally)

500

475
1

3
4
6
2
5
Pulses N until failure (log) [104]

7 8 9 10 11 12

Figure 6. Wöhler diagram

All tests concluded with the same mechanical failure. The damage pattern is shown in Figure 7. In the lower
picture, the crack formation has been visualized in red using a crack detection agent. Each failure started with afracture at the incomplete section of the threads which grew inwards after each subsequent impulse. Differences
between the different test specimens could be found regarding the consequences from the failure. For theconventionally manufactured valve block and the additively manufactured valve block without optimization, thefracture eventually reached the uppermost section of the threads and the fluid began to leak. For the additively
manufactured valve block with optimized geometry, the fracture experienced rapid growth and caused the fluid
to be sprayed from the leakage.

Crack

Figure 7. Damage pattern

CONCLUSION AND OUTLOOKThe project goal was to determine the differences between conventionally and additively manufactured valve
blocks and to conclude a statement on the suitability of additive manufacturing in hydraulics.
The experiments showed that an additively manufactured component has a lower stress tolerance than its
conventionally manufactured counterpart with identical geometry. This result is in accordance with the initial

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expectations. It is worth mentioning that exploiting the advantage of material saving by using an optimized
geometry further increases the risk of component failure.
Nonetheless, additive manufacturing can be advantageous. The pressures that were used in the tests were larger
than the pressures typically encountered in hydraulic systems. At commonly lower pressures, the difference
between conventional and additive manufacturing may potentially be neglected. Another advantage is thepossibility of post-processing which can eliminate cause of failure. Material usage can also be reduced by
employing a FEM analysis to identify points of mechanical failure. Given that all of the valve blocks exhibited
damage in the same location, this appears to be a promising option. Using additive manufacturing, it may bepossible to reinforce the critical locations and thus prevent crack propagation and enable a better fatigue
strength.
The next step is the closer examination of the afore-mentioned aspects. The experiment has shown that
additively manufactured components can withstand loads in the same order of magnitude as conventional
components. Further research may enable an improvement upon conventional manufacturing.

REFERENCES
1. DIN 8580: 2020, Manufacturing processes - Terms and definitions, division.
2. VDM METALS INTERNATIONAL GMBH: VDM Alloy 718: Nicrofer 5219Nb. https://www.vdmmetals.com/fileadmin/user_upload/Downloads/Data_Sheets/Datenblatt_VDM_Alloy_718.pdf, 02.04.2020.
3. SCHMID, M.: Additive Fertigung mit selektivem Lasersintern (SLS): Prozess- und Werkstoffüberblick, Springer
Vieweg, Wiesbaden, 2015.
4. LIEBHERR-AEROSPACE & TRANSPORTATION SAS: World Premier: First Ever 3D Printed Primary Flight Control
Component from Liebherr-Aerospace Flown on an Airbus Aircraft.
https://www.liebherr.com/shared/media/corporate/news/news-2017/04/13/aer/liebherr-airbus-first-flight-3d-flightcontrol-actuator-april2017.pdf, 01.04.2020.
5. LIEBHERR-AEROSPACE & TRANSPORTATION SAS: Liebherr-Aerospace Starts Serial Production of 3D Printed
Components. https://www.liebherr.com/shared/media/corporate/news/news-2019/01/17/liebherr-press-informationliebherr-start-3d-printed-series-production-jan2019.pdf, 01.04.2020.
6 Parker Hannifin Corporation: Slip-in Cartridge Valves: Pressure Control, Throttle, Check.
https://www.parker.com/literature/Hydraulic%20Valve%20Division/Catalogs/Catalog%20PDFs/Cat%20MSG143200_Slip-In%20Cartridge%20Valves.pdf, 11.05.2021.
7. GEBHARDT, A.; KESSLER, J.; THURN, L.: 3D-Drucken: Grundlagen und Anwendungen des Additive Manufacturing(AM), Hanser, München, 2016.
8. BUCHMAYR, B.; PANZL, G.: Eine SWOT-Analyse zum Einsatz der Additiven Fertigung für metallische Bauteile. In
BHM Berg- und Hüttenmännische Monatshefte, 2015, 160; S. 15–20.
9. LEWANDOWSKI, J. J.; SEIFI, M.: Metal Additive Manufacturing: A Review of Mechanical Properties: Annual
Rewiev of Materials Research; S. 151–186.
10. DIN EN 2624: 2008, Aerospace series - Pressure impulse testing of hydraulic system components.
11. DIN ISO 6772: 2018, Aerospace - Fluid systems - Impulse testing of hydraulic hose, tubing and fitting assemblies.
12. DIN EN ISO 19879: 2019, Metallic tube connections for fluid power and general use - Test methods for hydraulic fluid
power connections.
13. SUN HYDRAULICS CORPORATION: Data sheet: Model FAV Ninety degree Cavity: T-10A.
https://www.sunhydraulics.com/model/FAV, 01.04.2020.
14. DIN 51524-1: 2017, Pressure fluids - Hydraulic oils - Part 1: HL hydraulic oils, Minimum requirements.

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The Potential in Fluid Power Systems for a Sustainable Future
Katharina SCHMITZRWTH Aachen University, Institute for Fluid Power Drives and Systems (ifas)
Campus-Boulevard 30, 52074 Aachen, Germany(E-mail: Katharina.Schmitz@ifas.rwth-aachen.de)

Abstract. Climate change due to human made greenhouse gas emissions is the key challenge for the next decades.
Within the next years, the emission of climate active gases have to be reduced globally, requiring enormous changes
in all aspects of industry and daily life. In addition, all technologies need to get more sustainable, meaning not only
to allow for less greenhouse gas emissions but also respectfully treat all resources of our earth. Fluid power drives
and systems are an important technology in many industrial and mobile applications and have to take part in thistransformation process. In this paper, some potentials are pointed out to give a first impression on how fluid powersystems can help in developing a more sustainable future.
Keywords: Sustainability, Efficiency, Manufacturing, Tribology, Fluid Power Components, Mobile Systems

INTRODUCTION
Since more than three decades, researchers are aware of human made global warming,
induced by greenhouse gases in the atmosphere. With the first Kyoto Protocol adopted in
1997 and coming into effect in 2005 and 2012 respectively, first achievements to stop theever rising exhaust of greenhouses gases were made. In the Paris Agreement of 2016, mostnations agreed on the long-term goal to keep the rise in global average temperature below
1.5°C to max. 2°C. To reach this goal, the nations would have to reduce their greenhouse gas
emissions. Many ideas on how to achieve this ambitious goal in decreasing greenhouse gas
emissions near to zero until 2038 and latest 2050 exist. They are based on high penalties forindustry or prizing of CO2-emissions, the expansion of renewable energies like solar power,
wind power, wave and water power, more usage of hydrogen technologies, climate friendly
mobility and higher standards in the building sector for new buildings and many more.
Fluid power technology plays an important role in many industrial applications and branches.
Due to its high power density and its good controllability, the need for fluid power drives andsystems is unchanged. Therefore, the question must be asked and answered: What can fluid
power contribute to a more sustainable future? Some answers are given in the following
paper with, of course, no claim to be complete. In contrast, more ideas for a sustainable future
in general and in fluid power in detail are needed, requiring more creative engineering work.
POTENTIAL FOR SUSTAINABLITIY IN FLUID POWER COMPONENTS
When analyzing the potential for more sustainability in fluid power technology, it makes
sense to start at component level. Fluid power components are manufactured, leading togreenhouse gas emissions and the usage of other environment-impairing materials andprocesses.
The determination of emitted greenhouse gases in the production process is a complex
procedure with many unknowns. In a preliminary research project at ifas, the greenhouse gas
emissions during material production, component production, distribution, operation and endof-life were analyzed for pneumatic and hydraulic cylinders and translated in values for CO2
equivalents. Exemplary results of a pneumatic cylinder are shown in Fig. 1. One observationis that the operation of the exemplary cylinder, as well as in all the other examples, is
responsible for the largest amount of greenhouse gas emissions. Therefore, extra effort must
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be placed in energy efficient system designs and operation strategies of hydraulic andpneumatic systems as discussed later on.
Nevertheless, the emissions during material procurement and production can not be
neglected. Depending on the origin of the used material and the country of production, thevalues vary significantly. Therefore, measures have to always be based on a company andproduct individual analysis. However, it can be stated that there still is enormous potential forgreenhouse gas emission savings.

Figure 1. Greenhouse gas emission during the lifetime of a pneumatic cylinder, data given in
CO2 equivalents [see 1, 2]
One opportunity for a more sustainable manufacturing process is the use of digitalpossibilities in the manufacturing process. With a consistent use of data from development
and design stages to the use and end-of-use stage, development costs, reject costs, material
costs, transport emissions and energy can be saved [3].
A different challenge in hydraulic components are tribological contacts and the need forleaded material. Since lead is toxic, the reduction of these materials is important. Together
with partners from research and industry, ifas is researching new materials and coatings toallow the use of lead free material in hydraulic displacement machines. In Fig. 2 left, thelayout of a disc-on-disc tribometer is shown that is amongst others used for this research. On
the right side, exemplary measurement results are shown with different special alloys used forthe stator. The rotor is made of standard steel with a lapped surface. All materials aretechnical lead free special alloys with a fine grinded surface with exception of the blue curve,
OF2212. This material is a reference material with a lead content of about 0.4% and isregularly used in valves plates of displacement units.

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Figure 2. Disc-on-disc tribometer and Stribeck-curve of different special alloys (OF2212 is
the leaded reference material) [see 4]
From the measurements it can be seen, that several lead free special alloys have a goodpotential for an even lower coefficient of friction with comparable strain properties.
Additive manufacturing brings an additional potential to allow for new coating materials andcoating processes. One example is the EHLA-process (Extreme high speed laser welding) [5].
In addition, additive manufacturing can significantly help to reduce energy losses in fluid
power systems and components by optimized fluid flow inside the components in the future.
Another important role in a more sustainable fluid power is the used pressure media.
Pneumatic systems have the advantage of an environmental friendly fluid, but in hydraulicsystems, more effort must be taken to find alternatives to mineral oil. Pure water or water
based fluids can be part of the solution but definitely not for all applications. The same
applies to biodegrable oils. But when looking at greenhouse gas emissions of the fluid used, it
is worth thinking about alternatives. McManus, Hammong and Burrows presented a lifecycle assessment of mineral oil and rapeseed oil, see [6]. They came to the conclusion that forthe production of 1 kg of mineral oil 3.56 kg CO2 equivalent greenhouse gases are emitted.
For rapeseed oil only 0.3 kg applies.

POTENTIAL FOR SUSTAINABLITIY IN FLUID POWER SYSTEMSAs seen before, fluid power components bare diverse potential for more sustainableapproaches. When looking at systems level, two main pillars can be identified – measures formore energy efficiency in the fluid power transmission system and measures for an increasein productivity in the machine or application.
Energy Efficiency
Potential for more energy efficiency in fluid power systems can be found in various
examples, such as reduced throttling losses due to multi-pressure systems or displacement
controlled systems to enable recuperation, see [7, 8]. Often, energy saving measures are up to
now more expensive in aquisition compared to the saved cost in energy. However, when the3

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emitted greenhouse gases are taken into account, the reduction in systematic pressure losses
is crucial. The same applies for electrified machines like mobile construction machines. Bylowering the energy consumption of the hydraulic system, the required amount of energy
stored in an accumulator is significantly reduced and therewith costs for the electric
accumulator and installation space.
New power supplies of mobile construction machines additionally allow the integration ofelectric drives as well as the design of completely new hydraulic system architectures with
unprecedented degrees of freedom in size, rotational speed, control strategy, amount ofhydraulic pumps and therewith the whole hydraulic system. For a 9 t open-center compact
excavator, Opgenoorth presents possible system adaptations to improve energy efficiency,
see Fig. 3. By integrating variable motor speeds, variable pump displacements, electric valve
actuations and an electric swing, the energy efficiency of the excavator is increased by 42%.
[9]
System adaptations

Figure 3. Possible system adaptations based on an example of an excavator [9]In pneumatic systems, energetic losses are a controversial discussed matter. They exist, butsince energy is up to now, relatively cheap, cost consuming measures are not developed nor
produced. One promising new approach to save energy in pneumatic cylinder drive
applications without significantly increasing system cost, is the combined throttling. It
combines the advantage of the downstream throttling of load independent stationary speed
and the advantage of the upstream throttling of load dependent air demand. The challenge inthis new approach of ifas lies in the required binary control and the delayed start of motion
due to pressure adjustments in the downstream chamber.

Figure 4. Potential for more energy efficiency in pneumatic systems with combined
throttling [see 10]

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Increasing Productivity
The second pillar to more sustainability with regard to fluid power systems is the increase inproductivity. One example can be found at the construction site. In the last decades, theproductivity in all industrial sectors increased (world GDP +13% in last 20 years) due tomore automation. In contrast, in the construction sector productivity stagnated [11]. One key
for more productivity in this sector can be found in a higher automation – starting withautonomous functions of machines and ending with completely autonomous operated
construction machinery. In first, assisting systems like semi-autonomous functions have to be
fully developed and implemented. Challenging hereby are the design, accuracy and dynamics
of the controller. Potential solutions can be found in new controller design methods likeartificial intelligence methods or reinforcement learning methods [12].

Figure 5. Potential in assistance systems in excavators to increase productivity
For fully automated excavating functions, more research is needed and the integration ofmultiple sensors is required. With that, enormous potential exists for significant productivity
increase on construction sites and with that high potential for reducing greenhouse gas
emissions onside and in materials etc. Autonomous operating machines can also help
reducing the high number of human accidents and people being injured at work on
construction sides.
Another potential for fluid power systems to help the overall machine in operating more
efficiently is by the use of data from the fluid power system for condition monitoring
purposes. On one hand this allows , a good monitoring of the overall machine status and work
performance, and on the other hand the ability to predict machine or process failures before adamage or breakdown occurs [13]. One new approach for a better understanding of thesystem is by using so called soft sensors. In a new research project at ifas, a virtual volume
flow sensor is developed. Up to now, flow rate sensors are expensive and relatively difficult
to handle. The new approach is, to use two precise pressure sensors and a pipe in-between.
By the measurement of the dynamic pressures and a mathematical extension of the HagenPoiseuilles law for the transient time domain and turbulent flow, the volume flow rate can be
determined dynamically and with relatively low costs [14]. This virtual sensor allows a powermonitoring of a whole machine or individual actuators.

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SUMMARY
The preservation of the resources of nature and our earth for the next generations is one of themost important tasks for the next years and decades all over the world. By significantly
reducing the greenhouse gas emissions within the next years near to zero, global warming canhopefully be limited to less than 2°C. To achieve this goal, all drive technologies have toprovide the best effort they can to do so.
In this paper, several ideas and new approaches were presented to show a way how to
implement the sustainability aspect in the development of future fluid power systems andcomponents. Still, much more effort has to be placed in this field of research anddevelopment, leading to expedient solutions in the future.
ACKNOWLEDGMENTS
This paper shows work of several researchers at the Institute for Fluid Power Drives and Systems. The author
thanks the whole team at ifas, with special thanks to J. Duensing, S. Geffroy, C. Haas, A. Holzer, A.
Opgenoorth, O. Reinertz, and M. Rueckert for their contribution.

REFERENCES
1. MERKELBACH, S., Analysis of the Economic and Ecological Properties of Pneumatic Actuator Systems with
Pneumatic Transformers, PhD thesis, RWTH Aachen, 2019
2. DUENSING, Y., RICHERT, O., SCHMITZ, K., Berechnungstool zur Bestimmung des CO2-Fussabdrucks, O+P
Fluidtechnik, 2021/06, Vereinigte Fachverlage GmbH, Mainz, 2021
3. www.awk-aachen.com (accessed on 31 May, 2021)
4. HOLZER, A., SCHMITZ, K., Effizienzsteigerung von Verdrängereinheiten durch optimales Einlaufen, final report ofgovermental fundated research project, AiF No. 20083 N/1, 2021
5. HOLZER, A., et al., Tribological Investigations on Additively Manufactured Surfaces using Extreme High-Speed Laser
Material Deposition (EHLA) and Laser Powder Bed Fusion (LPBF), Proceedings of the 12th International Fluid Power
Conference, IFK 2020, Dresden
6. M. C. MCMANUS, G. P. HAMMOND, C. R. BURROWS, Life-Cycle Assessment of Mineral and Rapeseed Oil in
Mobile Hydraulic Systems, Journal of Industrial Ecology, Volume 7, Number 3-4, 2004, pp. 163-177
7. VUKOVIC, M., Hydraulic hybrid systems for excavators. PhD thesis, RWTH Aachen, 2017
8. WILLIAMSON, C., IVANTYSYNOVA, M., Power Optimization for Multi-Actuator Pump-Controlled Systems,
Proceedings of the 7th International Fluid Power Conference, IFK 2010, Aachen, pp. 91–102
9. OPGENOORTH, A., QUABECK, S., DONCKER, R. W. de; SCHMITZ, K., Challenges and possibilities of theintegration of electric drives in mobile machinery, Proceedings of the 12th International Fluid Power Conference, IFK
2020, Dresden, pp. 471–480
10. REINERTZ, O., SCHMITZ, K., Optimized Pneumatic Drives Through Combined Downstream and Adaptive Upstream
Throttling, Proceedings of the ASME/BATH Symposium on Fluid Power and Motion Control, 2020
11. MCKINSEY GLOBAL INSTITUTE, Reinventing construction: A route to higher productivity, 201712. DADHICH, S., BODIN, U., ANDERSSON, U., Key challenges in automation of earth-moving machines, Automation inConstruction 68, 2016
13. DUENSING, J., RODAS RIVAS, A., SCHMITZ, K., Machine Learning for failure mode detection in mobile machinery,
Proceedings of the 11. Kolloquium Mobilhydraulik, Karlsruhe Institute for Technology, 10. September 2020, pp. 1-26
14.https://www.ifas.rwth-aachen.de/cms/IFAS/Forschung/Digitalisierung-und-Automatisierung/~nqfoo/VirtuellerVolumenstrom-Sensor/?lidx=1 (accessed on 31 May, 2021)

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New Intelligent Hydraulic Power Control System
Takahiro URAI*, Ken SHINDO***

Sales Business Development, Marketing department, Bosch Rexroth Corporation
3-6-7 Shibuya, Shibuya-ku, Tokyo, 150-0002 Japan(E-mail: Takahiro.Urai@boschrexroth.co.jp)**
Engineering industrial Hydraulics, Sales department, Bosch Rexroth Corporation
3-6-7 Shibuya, Shibuya-ku, Tokyo, 150-0002 Japan(E-mail: Ken.Shindo@boschrexroth.co.jp)

Abstract. With the newly developed, intelligent hydraulic power control system, Rexroth combines the benefitsof hydraulics with low energy consumption and software-based functions. The integrated IoT service ensureshigher availability and avoids unplanned downtimes. With this IoT service, operators can monitor operating
statuses and plan maintenance in a cost-effective manner. As a result, fluid technology has taken a further step
towards the factory of the future.
Keywords: Industry 4.0, factory of the future, predictive maintenance, IoT

1. INTRODUCTION
The digitalization of mechanical and plant engineering means more than decentralized intelligence, integrated
sensors and connectivity. This applies to hydraulics too. With its power density and immunity to impacts andvibrations, hydraulics plays a central role as an important drive technology in Industry 4.0 too. The compact
intelligent hydraulic unit is revolutionizing hydraulic pressure supply. This unit is a modular hydraulic power
unit for the medium performance range from 7.5 kW to 30 kW. It comprises optimally tailored, impressively
compact components. Compared to previous units, considerable amounts of space can be saved.

2. PERFORMANCE DESCRIPTION
One of this system’s unique selling points is it compactness. This is possible because the hydraulic power unit
can be integrated into a factory like a control cabinet, i.e. upright. The tank no longer supports all othercomponents – Rexroth developed a special supporting structure for this purpose. In addition, a degassing flowoptimized tank is used. On the basis of extensive CFD simulations, the developers improved the tank geometry
so skillfully that the new unit needs only a quarter of the oil required by classic designs: instead of 600 liters,
150 liters are now enough – with at least the same service life. The drive unit – a component which usually takes
up a large amount of space in hydraulic power units – is equally compact. The asynchronous servo motor has
been replaced by a synchronous servo motor. This is the most efficient motor design, producing the same torqueeven though it is 80 to 90% smaller. The manifold too was redesigned: thanks to 3D printed sand cores, theblock including the internal flow geometry is now cast. As a result of this flow-optimized inner structure, lessmaterial and space are required. Thanks to all these measures, this unit has a footprint of just 0.5 square metersinstead of one and a half to two square meters（Fig.1）. This represents a saving of up to 75%. On top of this,
there are energy savings of up to 80% as well as a 10 dB(A) reduction in noise emissions. As a result, noise
emissions are lower than 75 dB(A). Refer Table.1 for technical feature..

3. CONNECTED FOR THE FACTORY OF THE FUTURE
In this unit, machine manufacturers will find a ready-configured drive controller with numerous integratedfunctions. Various sensors provide information on the current filter, oil or drive status. The collected sensor data
are bundled via IO-LinkMaster and pre-processed by the drive controller so that they can be integrated flexibly
into modern machine designs. This means that this intelligent unit is IoT-ready.

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Figure 1. Dimensions
Table 1. Technical feature

Pressure
Hydraulic power
Flow rate
Oil volume
Drive
Cooling system
Cooling water
Noise

Up to 31.5MPa
30kW
Up to 160L/min
150L
Servo motor, watercooled
Plate heat exchanger
20L/min<75dB(A)

4. IoT SYSTEM MONITOR SERVICE
With the IoT digital service, operators have all information regarding this system at their fingertips at all times –whether it be information regarding the standardized visualization of the component and operating status or
chargeable IoT services such as forthcoming maintenance work and predictive maintenance analyses using
Rexroth’s Online Diagnostics Network (ODiN).
The data collected are transferred to the browser-based Monitor web dashboard via Multi-Ethernet or 4G-LTE.
As a result, operators can be kept up to date regarding the current operating status and key status indicators on
any end device (tablet, smartphone, PC). This plug and play service is free of charge and can be used without
additional installation work. This IoT monitor service system is implemented in this unit as standard.

Figure2. IoT system monitor example

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5. ADDITIONAL ANALYSIS FUNCTIONSIn addition to this unit’s automatic status monitoring, maintenance personnel and maintenance managers can add
extra solutions for various applications as add-ons. These pay-per-use payment models include additional IoT
analysis tools and can be subscribed to on a monthly basis.
For optimizing maintenance processes, Maintain module in this IoT system monitor service offers access tohistorical sensor data as well as messages if maintenance is required or in the event of critical operating statuses.
Pre-defined and custom rules allow customers to improve their maintenance strategy on an ongoing basis.
Customers benefit from scalable software solutions from the Bosch IoT portfolio such as the Nexeed Production
Performance Manager and the integrated know-how of the Rexroth domain specialists. The Nexeed Production
Performance Manager is a condition monitoring software solution for systematic production optimization.
With Maintain module, data concerning the most important components for reliability and operating life (thehydraulic oil and the drive unit) are assessed. Comprehensive insights into the correlation between the speed of
the servo motor and the flow allow the monitoring of system leakage, while correlations between motor data andthe operating pressure allow conclusions regarding drive behavior. The leakage sensor, temperature and levelsensor, particle sensor, water sensor and oxygen sensor are additional data sources which can be used to make astatus diagnosis.

Figure3. Maintain module example

6. PLANNING PREDICTIVE MAINTENANCE EFFIFIENTLY
A further add-on, in Predict module, uses the machine learning algorithms of the Rexroth Online Diagnostics
Network (ODiN) for a predictive analysis of the system. With this IoT service, the power curves for a wide
range of components are recorded for data analysis purposes. In the event of deviations, the system
automatically calculates the expected remaining operating life of the relevant component and informs theoperator immediately via a push message. As a result, predictive maintenance application cases can be taken
into account predictively in the operator’s maintenance plans, thus ensuring maximum availability.

7. CONCLUTION
This new form of hydraulics is energy-efficient, quiet, space-saving and intelligent. With the integrated IoT
system monitor service, this intelligent unit offers anything from transparent, efficient machine monitoring tointelligent predictive maintenance measures. Thanks to this unit, fluid technology is now one step closer to thefactory of the future.

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