Special Issue;
Automotive Technologies
November 2011ISSN 0915-0528 OSAKA, JAPANR

TECHNICAL REVIEW
No.79

Special Issue:
Automotive Technologies

corporation

NTN TECHNICAL REVIEW No.79
CONTENTS
Yoshikazu FUKUMURA

Preface
Contribution
Perspective

Prospects for Environmental and Energy-related Vehicle Technologies
Yasuhiro DAISHO Professor, Waseda University Graduate School of Environment and Energy Engineering

Market and Technology Trends in the Automotive Industry
Makoto OKASAKA

1

2

12

Special Issue for Electric Vehicle ● Earth-friendly NTN - Pioneering Technologies for Tomorrow
In-Wheel Motor System

22

Yuichi ITOH, Kayo SAKAI and Yusuke MAKINO

Development of In-Wheel Motor System for Electric Commuters
Tetsuya YAMAMOTO, Aiko ISHIKAWA and Wataru YAMADAOne Motor Type Electric Vehicle Drive System

29
33

Fumihiro ISOBE, Yoshinori ITAKURA, Yusuke OOSUMI and Guodong LI

Development of a New Steer-by-wire System

42

Katsutoshi MOGI, Tomohiro SUGAI, Ryo SAKURAI and Nobuyuki SUZUKI

Development of Two-seat Electric Vehicles with In-Wheel Motors
Akira YAMAGATA, Aiko ISHIKAWA, Chinami ITOU, Tomomi GOTOU, Kayo SAKAI, Yukiko TAKEKAWA,
Kaori TOMODA and Daisuke MATSUOKA

51● Automotive Products and Engineering
Hub Bearing with Integrated Multi-axis Load Sensor

58

Kentaro NISHIKAWA

Hub Bearing for Severe Environments

64

Yasushi SHIBATA

Lightweight and High-efficiency Drive Shafts for Rear-wheel-drive Cars
Tatsuro SUGIYAMA and Yuichi ASANO

Ball Screw Unit for Automotive Electro-actuation

69
73

Koji TATEISHI

Low-torque Deep-groove Ball Bearings for Transmissions
Katsuaki SASAKI

Technology Trends in Auto Tensioners

78
83

Seiji SATO

NTN-SNR Low-torque Strut Bearing for Severe Environments
Gérald MIRABEL and Susumu YAMAJI

NTN-SNR Low-torque High-capacity Pulley Bearing Unit for Water Pumps

90
94

Sebastien BRISSON

Development of a System for Rolling Bearing Design Optimization
Daisuke IMADA, Tsuyoshi NIWA, Takashi UENO and Tomohisa UOZUMI98● Technical Papers Technical Articles New Products
Rapid Evaluation of Shear Fatigue Properties of Rolling Bearing Steels for Lifespans Up to the Gigacycle Range
Noriaki SAKANAKA, Yukio MATSUBARA, Yoshinobu SHIMAMURA and Hitoshi ISHII

Two-Dimensional Dynamic Analysis of Cage Stress for Needle Roller Bearings
that Support Connecting Rods in Reciprocating Engines
Tomoya SAKAGUCHI
Integrated Bearing Dynamic Analysis System (IBDAS)
Mariko SEKIYA

Hybrid PEEK Sliding Bearing

104
111
119
125

Takuya ISHII and Ken YASUDAAir Spindle for Ultra High-precision Machine Tools

130

Teruyoshi HORIUCHI and Kazuyuki AONO

● Our Line of Award Winning Products”2010 ‘CHO’ MONODZUKURI Innovative Components Awards” Automotive Component Award
Torque Diode for Seat Lifter
Masahiro KAWAIThe First Approval of Tribology Inheritance

Journal Rolling Bearing for the 0 “Zero” Series Shinkansen Trains
Our Line of New Products

Kouya OHIRA and Masanori UENO

136
137

138

NTN TECHNICAL REVIEW No.79（2011）[ Preface ]Introduction to this special issue
on automotive technologies

Yoshikazu FUKUMURAManaging Director

As measures to reduce greenhouse gases and balance power supplies and demands in theenergy field, the utilization of renewable energy sources, including the generation of power withwind and solar, has been accelerating. In the automotive field, the development of nextgeneration automobiles, including hybrid cars and electric vehicles, has been advancing.
In this changing environment, we are further improving the technologies that we have cultivatedthus far at NTN, and we are developing green products based on the themes of reducing theirtorque, increasing their life spans, and making them more compact and lightweight in order toincrease the conservation of resources and energy. Furthermore, we established our EV System
Division in April 2011 to accelerate the development and commercialization of drive systems,
steering systems, and control systems for electric cars, one type of the next generation ofvehicles.
We have worked to publish this issue of our Technical Review, which features technologies andproducts for cutting-edge automobiles, in time for the 42nd Tokyo Motor Show that will be held
over 10 days, December 2–11, with the theme “Mobility can change the world.” This issue starts
with “Prospects for Environmental and Energy-related Vehicle Technologies,” a contribution byProfessor Yasuhiro Daisho who is an authority in the automobile environment and energy field at
the Waseda University Graduate School of Environment and Energy Engineering. Following this,in our Special Feature on Electric Vehicle Technologies, we introduce our in-wheel motor system,
one-motor-type drive system, and steer-by-wire system for electric vehicles. We have been
conducting research and development for these systems from standpoints that are completely
different from the past.
At NTN, we will celebrate our 100th anniversary in fiscal 2017. Following our theme, “For New
Technology Network: We shall contribute to international society through creating newtechnologies and developing new products,” we are striving to expand our business in the three
years from fiscal 2011. To do this, we are implementing our Global Advance 2013 medium-term
management plan, in which we stand on a foundation that puts “technology first” and we seek torealize a strong corporate composition that does not rely on scale. With this plan, we are also
promoting the development of products that are better for the environment and contribute to thesustainable advancement of society.

-1-

NTN TECHNICAL REVIEW No.79（2011）[ Contribution ]Prospects for Environmental and Energy-related Vehicle Technologies
Improving Conventional Technologies and Electrifying Vehicles

Yasuhiro DAISHOProfessor, Faculty of Science and Engineering, Waseda University

To comply with increasingly stringent emission regulations, automakers are being forced
to develop ultra-low-emission engine systems by optimizing combinations of technologies
related to combustion, exhaust after-treatment and fuels. These vehicles are expected toretain their importance with state-of-the-art technologies for two decades to come. Moreemphasis will be placed on improving fuel economy and promoting the popularity of hybrid
and electric vehicles, decreasing vehicle weights, and utilizing renewable energy and fuels
to reduce oil dependency in the transportation sector, thereby mitigating global warming.
High-performance and cost-effective components are essential for these electrified
vehicles. While keeping an eye on these trends in the global market, we should develop acomprehensive technology strategy.

expected to increase 40% by 2030 if current
measures are left unchanged. Above all, 60% of
petroleum use will be in the transportation field,
primarily caused by the remarkable increase indemand for petroleum in emerging nations, inparticular, China, India and countries in South East
Asia, where motorization is progressing
considerably.2)
Under these circumstances, the development oftechnologies that improve fuel efficiency even more
and the diversification of energy sources havebecome indispensable. Automobile manufacturers,
betting on their own survival in the international
market, are working on the development oftechnologies related to these things. In this paper, I
will examine the prospects for automotive
technological problems related to the environment andenergy and their solutions from a medium and longterm perspective, while offering my own opinions.

1. Introduction
In the industrialized nations, not only are entire-car
manufacturers continuing to develop andcommercialize novel technologies but the various
automotive material and component manufactures
have formed a big industrial field for commercializingthese technologies. The common use of automobiles,
as a means of mobility and transportation, has brought
convenience and abundance to our lifestyles. On theother hand, automobiles consume a large amount of
petroleum, and are a main source of emissions thatcause urban air pollution and CO2, which are typicallyconsidered greenhouse gas. We can expect, however,that by the middle of this decade, vehicles ranging
from passenger cars to heavy trucks will becomecompliant with strengthening restrictions on exhaustemissions, and the problem of air pollution will havebeen largely overcome in the industrialized nations.1)
To address global warming, Japan, in fiscal 2008
began a five-year effort with the goal of reducing
greenhouse gases by 6% when compared to fiscal
1990 levels to meet the mandate of the Kyoto
Protocol. In 2013, each country must set forth newtarget values for reducing greenhouse gases by 2020.
Currently, 17% of the total CO2 emissions in Japanare from automobiles, and measures to reduce them
are becoming an even more serious concern.
According to the World Energy Outlook 2010, which
is a report by the International Energy Agency, theprimary energy demand of the entire planet is

2. Improvement in Conventional Vehicles
2.1 Fuel economy improvements for gasolinepowered automobiles
For gasoline-powered automobiles, emission
controls against the three pollutants—NOx,hydrocarbons (HC) and CO—are advancing greatly
through the combination of more accurate
electronically-controlled fuel injection systems andthree-way catalytic converter systems. In Japan,
reduced taxes on green cars and purchase subsidy-2-Prospects for Environmental and Energy-related Vehicle Technologies

Table 1 Fuel economy improvement measures and theirratios for the proposed 2020 Japanese passenger car
fuel economy standards

programs are succeeding, so most gasoline
automobiles are now super-low emission models that
emit one quarter of the regulated upper limits of NOxand HC, and impacts on the atmosphere have been
greatly curtailed. While it can be said that this is atrend shared by other industrialized nations, thesituation is serious in emerging nations, which haveair pollution caused by traffic congestion and theresulting automobile emissions in large cities. Stronger
regulations to improve fuel characteristics including
reductions in sulfur content as well as theincorporation of sophisticated emissions control
technologies are now thought to be necessary.
From now on, in addition to maintaining low exhaustemissions, improving fuel economy is becoming aneven more critical issue for gasoline-powered cars. In
Japan, the fiscal 2010 fuel economy standard3) was
achieved ahead of schedule, and a standard for fiscal
2015 has been proposed that seeks fuel economy
increases that are 23.5% better than that of fiscal
2004.4) Furthermore, Corporate Average Fuel
Efficiency (CAFE) standards are being applied toward
strengthening the fiscal 2020 standards for passenger
cars, and considering the popularization of hybridcars, a proposal has been made to seek an averageof 20.3 km/L, which is a 24.1% improvement from thecomposition of vehicles in fiscal 2009. The plan toincrease fuel economy numbers for hybrid vehicles will
be determined next year.5) Specifically, after
determining fuel economy target values for eachvehicle weight class, what is sought for each company
is that the weighted harmonic mean value from theactual fuel economy values for the number of vehicles
of each class shipped does not go below the weighted
harmonic mean value for the number of vehicles
shipped according to the fuel economy target values
set for each class.
In the EU, corporate average fuel efficiency
standards based on CO2 emission amounts havebeen set, and from the 130 g/km target value for 2012
to 2015, 95 g/km has been proposed for 2020. Even in
the USA, as though chasing after Japan and Europe,stronger standards for 2017 to 2025 have been
proposed.
Specific examples of technologies to achieve thesetargets, including fuel economy improvement
measures that were considered during investigationsof the above fiscal 2020 fuel economy standards, andtheir improvement ratios are shown in Table 1.5) They
include the use of a variety of advanced engine
variable geometries, direct injection and other fuelsupply system control refinements, CVT, DCT andother more efficient transmission systems, engine size
reduction through the use of supercharging systems,

Fuel economy
improvement
ratio
Friction reduction
1%
4 valve
1%
2 valve + 2-point ignition
2%
Variable valve actuation
1〜6%
Electromagnetic valve actuation
10%
Direct injection engine
2〜10%
Variable cylinders
7%
Miller cycle
6%High-volume exhaust gas recirculation (EGR)
2%Heat management (cooling loss reduction, heat recapture, etc.)2%
Switching to variable compression
10%
Supercharging size reduction
8%
Electric power steering
2%Electrification (electric-powered wipers etc.)
1%
Recharging control
0.5%
Idling neutral control
1%
Increasing automatic transmission (AT) stages
2%
Expanding AT lockup range
2%
Continuously variable transmission (CVT)
7%

OthersRunning
resistance
reduction

Improvements
related to drive

Auxiliaryequipment
loss
reduction

Engine improvements

Fuel economy improvement measure

Semi automatic transmission (AMT),
Dual clutch transmission (DCT)

9%Manual transmission (MT)9%Rolling resistance reduction

1%

Aerodynamic improvements

1%

No idling (other than hybrid automobiles)
Diesel vehicles

7%20%

No idling + energy recovery(other than hybrid automobiles)

10%

and the reduction of mechanical friction in every partas well as the reduction of parasitic losses fromauxiliary systems. Furthermore, as I will discuss later,
hybridization and vehicle weight reduction are also
effective, and these must be integrated into seeking
large overall fuel economy improvements. Until now,
fuel economy improvements of 10 to 20% or so havebeen achieved. For engine technologies themselves,
we can predict that, having meeting the strengthened
standards of fiscal 2020, we will enter a period of
decreasing returns as we approach an improvementof about 20%.

2.2 Countermeasures for exhaust gas and new
combustion methods for diesel vehicles 6)
Diesel engines have good fuel economy, and are atype of motor that will continue to be popular in the
future for trucks and buses, which require high output
and durability. On the other hand, NOx, black smoke
and particulate matter (PM) are emitted at the sametime due to combustion of non-uniform fuel spray, soemissions that are as clean as those from gasoline-3-

NTN TECHNICAL REVIEW No.79（2011）

Motors have introduced clean diesel passenger carsthat conform to the Post New Long Term
Regulations.1) For Japan, from the perspective of
product balances in petroleum refining and CO2
reduction, there are expectations for the revival ofdiesel cars, but the keys to their popularization on abroad basis are the development of NOx catalytic
reduction with high purification capability and the cost
reduction for the overall exhaust system.7)

vehicles are sought in Japan, America and Europe.
Last year, in accordance with the world harmonized
test cycle (WHTC), regulated values for emission gas
testing methods for NOx and PM to be implemented in2016 were proposed at 0.40 and 0.01 g/kWh,respectively, in Japan.1)
As shown in Fig. 1, NOx countermeasures include
exhaust gas recirculation (EGR) and injection timing
control. Measures to improve fuel economy and PM
include the utilization of variable compression
mechanisms and multistage turbocharging systems,
as well as common rail systems capable of highpressure, and flexible multistage fuel injection thanks
to electronic control. We predict that supercharging
will improve emissions even more, allowing engine
size to be reduced further. Departing from the present
fuel injection pressure of around 200 MPa, it ispossible that there will be demand for pressure ofnearly 300 MPa, and it will also be necessary toaddress demands for greater rigidity, reliability anddurability in related parts.
Furthermore, as emission after-treatment technologies,
in addition to diesel particulate filters, it is necessary to
combine urea-based selective catalytic reduction(SCR) and NOx storage catalytic reduction. It is also
necessary to ensure reliability and durability as well as
the clear-cut division of roles with combustion
technologies. Issues related to managing the costs of
the overall system must also be overcome. We canpredict that eventually these types of technologies will
converge into optimal systems that are common to
any manufacturer.
Moreover, although diesel passenger cars, whichhave achieved high performance, comprise about
50% of all passenger cars in the EU, due to smokeand noise issues, they have been held back in Japan,
and they had disappeared from the market with theincreasing costs of exhaust emissions control.
Recently, however, Nissan Motors and Mitsubishi

EGR
coolerAir filter
Slot valve
E

Oxidizing catalyst + DPF+ NOx reducing catalyst

Fresh airE
Turbosupercharger(variable-pressure
mechanism, 2-stage)
Intercooler

Air intake
slot valve

EGR valve
E
E

E
Reducing
agent supply
system
EGR
cooler

3. Electrification of Automobiles

8)

3.1 History and background of electrification
After the oil shock that broke out in 1973, there wasinterest in electric vehicles (EV), but development did
not advance, in part because of the eventual decline
in the price of crude oil. Then, in the 1990s, the US
state of California enacted a low-pollution vehicle
program. As one part of this, it was proposed that one
in every ten passenger cars sold in the state would be
required to be a zero emission vehicle (ZEV).9) EVs
are really the only kind of ZEVs, so a development
boom occurred, but they were not well-received byconsumers, and the program was revised to allow
gasoline vehicles with low-polluting exhaustemissions. As a result, EVs did not achievemainstream popularity. In every case, the batteries,
which are the key components, were heavy and bulky,
required a long time to recharge and only allowed
short travel distances, not to mention being expensive.
These drawbacks prevented their popularization.
Then, starting in the 1990s in the United States, thePartnership for a New Generation of Vehicles (PNGV)
was started as a national project to triple milage10) The
big three American automakers worked on
hybridization as the main technology, but they did notachieve the goal on a commercially acceptable level.
Prompted by this, however, Japanese auto
manufacturers created hybrid cars that were ready for
consumer use by the late 1990s. Furthermore, EVs
that carry higher performance batteries, as well as
fuel-cell vehicles (FCV) that use hydrogen as fuelhave also appeared recently. Through this process, as
shown in Fig. 2, research and development advanced
greatly in areas ranging from batteries and motors to
inverters, DC-DC converters and other electronic
devices, as well as in power electronics, including
power-energy management, vehicle weight reduction,
high-efficiency engines for hybrids and fuel cellsystems. Japan currently has a great lead in thesetechnological fields. I will discuss it later, but as an
example, in Fig. 3, I provide an overview of the electric
system of the Mitsubishi Motors i-MiEV, which is thefirst EV in the world to be a truly mass-produced.

Slot valve
E
EmissionsTemperature-pressure
inputs from various parts
Control
outputs
to various
parts

ECU
E
ElectronicallyE : Inputs/outputs for electronic
controlledcontrol
high-pressure
injection system(multistage injection)

Fig.1 Typical example of measures to meet future diesel
emission regulations-4-

Prospects for Environmental and Energy-related Vehicle Technologies

3.2 Hybrid vehicles 8, 11)

I expect that the determining factors for morecommon use of electric vehicles will depend on
performance of lithium-ion (Li-ion) batteries in terms ofboth energy density and power density: more common
use of Li-ion batteries will be promoted by further
improvement in efficiency and lower cost.
Furthermore, the most common electric motors usedin vehicles are high output variants of highly efficientwater-cooled permanent magnet AC synchronizedtype motors using neodymium and dysprosium. Longterm availability of raw materials of these rare earths
will pose another challenge.

Currently, as indicated in the background above,
hybrid technologies that combine engines, motors,
generators and batteries are the most promising oftechnologies that greatly improves fuel economy
compared to conventional engine-powered cars.
Types of hybrid systems can be classified as shown
below, and the structures of types [2] and [3] areillustrated in Fig. 4. Toyota and Honda are alreadymass-producing such vehicles and cost reductions arein progress. At this time, their popularization is aheadof EVs, which still have high costs.
[1] Micro hybrids: The engine is started and stopped
by a motor (stop idling) and regenerative braking
occurs during deceleration to recharge. With thesefunctions, fuel economy is increased by about
5–15%. The motor may be incorporated into thetransmission system and used as a starteralternator, leading to size reduction of the vehicle.
[2] Mild hybrids: In addition to the functions of micro
type hybrids, these have systems to provide power
assistance (parallel type). Fuel economy
improvement is about 20–50%. Examples include
Honda Motor Company’s Insight, Fit and CR-Z,
which all use Ni-H batteries, and Nissan Motor
Company’s FUGA hybrid, which uses lithium
batteries.
[3] Full hybrids: This type of hybrid includes a motor
and power generator. There are series types
whose engines are used only to generate electric
power, and dual types, which have both series andparallel functions. Fuel economy improvement
ranges from about 50 to 100%. Examples of dual
type hybrids are Toyota Motor Corporation’s Priusand SAI, which are series/parallel types that use
Ni-H batteries.

Small and medium-sized EV-1970s, 1990sDevelopment ofelemental technologies
Motors Batteries Inverters DC-DC converters
Various electronic devices Electronic control
Engines Weight reduction

Small andmedium-sized EV

Fuel-cell vehicles
Hybrid cars

Fig. 2 Future trends in the electrification of vehicles

User’s recharging lid

Drive battery

High-voltage connector

Inverter

Boosting
recharging lid

Onboard recharger& DC-DC converter

HeaterMotor (power unit)
Service plug
High-voltage connector

AC compressor

Fig. 3 Electric drive system of i-MiEV(Source: Mitsubishi Motors Corporation)★Micro hybrids have stop idling and regeneration functions but do not have power assistance functions.
Fuel economy
System
M: motor
G: generator
C/I: controller/inverter
improvement rate

B: battery
T: transmission system
Ps: power split device
Pi: plug-in：Power/generation

E

B

C

M/G

T

C

C/I

M/G

G

C/I
Pi

20〜50%

E

Ps

B

C/I

G
B

Pi

Parallel (mild)E

C: clutch
E: engine：Regeneration

Series (full)50〜100%Fig. 4 Various hybrid systems

-5-

M/G

Pi

Series/parallel (full)50〜100%NTN TECHNICAL REVIEW No.79（2011）

Gasoline hybrid vehicles almost double fuel economy
at best, exceeding the fuel economy of diesel cars. For
still better fuel economy, diesel hybrid cars must be
developed. However, due to increased costs for diesel
passenger cars, that include exhaust emissions
controls, we can predict that achieving
commercialization will not be easy. At the same time, in
response to future strict CO2 restrictions in the EU, as
described above, the trend to apply parallel hybrid
systems particularly with high-class gasoline cars anddiesel cars, has increased in the last 2 or 3 years.
Moreover, various companies have been releasing
diesel trucks for regional transportation and highway
buses that are parallel hybrid vehicles in Japan. These
achieve fuel economy improvements of 20 to 30%, but
because of the large size of the vehicles, further
improvements, including size reduction of hybrid
system and batteries, are necessary.

Features of these overall systems include the merits
that the motors have high-efficiency in a wide
operation range and that the battery loss during
recharging and discharging is low. In addition, EVs
and hybrid vehicles use their motors as generators togenerate electricity while braking and store it in theirbatteries, allowing them to recover part of theiroperation energy during deceleration. By doing this, asignificant energy conservation effect is achieved.
This type of braking is called “regenerative braking”
and is a method that is applied in conjunction withmechanical brakes.
Following the beginning of sales of Mitsubishi Motor
Company’s i-MiEV cars, which carry Li-ion batteries, in
2009, Nissan Motor Company’s Leaf appeared in themarket as a mass-produced EV at the end of 2010.
Electric power has the great merit of being
significantly less expensive than gasoline, considering
the practice of discounted electric power rates atnight, for example. On the other hand, since thebatteries are expensive, they are still fairly expensive
vehicles at present. However, Toyota and Honda will
also enter the EV market in 2012 with models basedon the iQ, RAV4 and Fit.
Furthermore, compared to conventional engine
vehicles, the number of parts in EVs is 20 to 30% less,and a manufacturing approach with a so-calledhorizontal specialization of work is possible bycombining major types of parts in modules. In fact,
venture businesses in the United States have started
manufacture and sales using this type of approach. In
addition, while on a small-scale, sales of simple
compact cars for short distance travel with limited
speed and modification kits for gasoline cars havebegun. Assuring safety for these vehicles, however,seems to be a challenge. Existing automobile
manufacturers can be said to have advantages when
they decide to mass-produce vehicles, which allows
costs to be reduced, and in establishing networks that
include sales, repair and maintenance. Anyway welook at it, however, we can see that a diverse varietyof EV types will appear that respond to the different
needs of a wide range of regions and income levels in
the future.

3.3 Electric vehicles8)
As discussed above, since exhaust cleaning
technologies for conventional vehicles have greatly
advanced, for example, the necessity of the zero
emission feature of EVs has declined, and interest has
been focused on their CO2 emission reduction andenergy conservation features. Considering the current
power source composition, Well-To-Wheel (overall
performance from the production of fuel from primaryresources and energy to transportation, storage andvehicle operation) energy efficiency and CO2
emissions are both far superior to other types ofvehicles. This is demonstrated clearly in results ofcomparative evaluations of the current Well-to-Wheel
efficiency and CO2 emissions of various types ofautomobiles provided by the Japan Hydrogen & Fuel
Cell Demonstration Project, as shown in Fig. 5.12)

Total CO2 emissions per km traveled (10-15 mode)
Vehicle type

Unit: g-CO2/km

0

50

100

150

200

Current FCV
Future FCV
Gasoline
Gasoline HV
Diesel

3.4 Plug-in hybrid vehicles

Diesel HV

Recently, a type of automobile called a “plug-in hybrid
vehicle (PHEV)” has begun to enter the mainstream
both in Japan and abroad. These vehicles feature the
combination of both a charge depleting (CD) mode of
travel that uses electric power charged from an externalsource and a charge sustaining (CS) hybrid travel mode.
With this combination, these vehicles benefit from boththe low CO2 emission advantage of the power source

CNG
BEV (Battery EV)
Current FCV: Calculated using top values from JHFC empirical results for hydrogenstation and FCV data and top values from literature for other data
Future FCV: Calculated using a Future FC system efficiency of 60% for FCV and
top values from literaturePower source composition: Average power supply composition for Japan

Fig. 5 Comparison of CO2 emissions in various vehicles(Source: JHFC, March 2006)-6-

Prospects for Environmental and Energy-related Vehicle Technologies

Furthermore, in order to realize their widespread
adoption, along with manufacturer efforts to reduce
costs, programs that respond to their current highcosts, including reduced national taxes and purchasesubsidies to cover the differences in prices overordinary vehicles, the establishment of recharging
stations and greater common awareness about theiruse, which is limited to short and medium distances,
are necessary. Overseas, plans are being made topopularize these vehicles with levels of one million
vehicles by 2015 in the United States and two million
and one million vehicles by 2020 in France andGermany, respectively. In all these countries, we canpredict that these five and ten year points will be
critical to achieving widespread adoption.
In the United States in particular, since the change
to the Obama administration, as part of an energy
policy to reduce dependence on petroleum, plans arebeing implemented to construct a smart grid that willuse IT to network conventional electric power sources
with renewable energy sources, and policies are being
advanced to support the popularization of EVs and
PHEVs that will utilize this grid.
As a support policy for these types of vehicles inJapan, the Ministry of Economy, Trade and Industry
designated 11 cities as “EV & pHV Towns” in fiscal
2009 with the goal of achieving a total of 3,200
vehicles and installing 10,000 boosting and regular
recharging equipment units. These designations weremade in order to take the lead in EV and PHEV (alsoabbreviated PHV) adoption and work on how they areused and resolve related problems. Since then,additional designations have continued to be made.15)
Furthermore, through support from the Agency for
Natural Resources and Energy, the Committee for
Next Generation Energy and Social Systems wasstarted in 2010 with the purpose of investigating thepotential for the effective use of smart grids, including
the recharging of vehicles. Yokohama, Nagoya, Kyoto
Keihanna Science City and Kita Kyushu City havebeen designated for their investigations, and empiricaltesting has begun in these four regions.16) As an
example, let me introduce a smart house proposal that
Toyota is working on. As shown in Fig. 6, it includes
installation of a home energy management system(HEMS) that can manage energy use by PHEVs and
EVs as well as within the residence. This systemenables the comprehensive management andregulation of overall supply and demand, including
electricity from power companies and energy
generated by the household from solar and othersources. In addition, this system provides information
to residents and vehicle users and allows themexternal control.

itself and the low cost of recharging, as well as theability to travel long distances in CS mode. This canbe said to offer the benefit of eliminating driverconcern about running out of battery energy. In 2009,
the Ministry of Land, Infrastructure and Transport
presented a method to measure exhaust emissions
and fuel economy for this type of vehicle.13)
Toyota Motor Corporation began limited sales of
600 Prius Plug-ins, its first PHEV, in 2010 and isconducting demonstration tests. Using a Li-ion battery,
this model is capable of traveling 23.4 km as an EV.
General sales are scheduled to begin in 2012.
Another example is the mass-market sale of the
Chevy Volt begun by GM at the end of 2010. In
addition, Honda and Suzuki also plan to start sales of
PHEVs. Honda plans to release a model based on the
Inspire in 2012 and Suzuki plans to release a model
based on the Swift in 2013.
Obviously, if the size of the battery in a PHEV isincreased, the travel distance in CD mode can beextended, but this invites increased costs and greater
energy consumption due to the increased weight.
Further improving the performance of the batteries
and reducing costs are challenges, but compared toEV, they already have both cost and convenience on
their side when it comes to potential popularization. In
any case, it is necessary to estimate the average daily
travel distance of ordinary drivers and determine theoptimal amount of batteries to be included onboard.

3.5 Measures to support EV and PHEV popularization
The Ministry of Economy, Trade and Industry has set
target values for improving the performance of batteries
for EVs and PHEVs and lowering their costs, as shownin Table 2.14) In particular, improving energy density,
which leads to weight reduction, is the most important
issue. Research and development support is also being
provided through NEDO and other organizations.

Table 2 Targets for developing Li-ion batteries for electricvehicles (Source: NEDO and METI, 2006)
Phase

Application
Performance

Improvement Advancement
2010
2015

Compact EVs EVs and HVs
for electric
for limited
power
commutingcompanies

EVs, FCVs
and plug-in
HVs forcommuting2020?

Innovation
2030

Highperformanceplug-in HVs

Widespread
adoption of
EVs

1

1

1.5

3

7

Energy density

100Wh/kg
400W/kg

100
1,000

150
1,200

―
―

700
1,000

Energy density

70Wh/kg
1,900W/kg

70
2,000

100
2,000

200
2,500

―
―

1
(20)

1/2
(10)

1/7
(3)

1/10
(2)

1/40
(0.5)

EV Output density
HV

Present

Output density

Cost(¥10,000/kWh)Cooperation

Development Led by private Led by private among industry,
government
structure
sectorsectorand academia

Universities
Researchorganizations-7-

NTN TECHNICAL REVIEW No.79（2011）

oil, including used oil, and other oils that have been
converted to methyl ester. In Japan their use inalready existing vehicles is permitted by the Act on the
Quality Control of Gasoline and Other Fuels if mixed
in concentrations of 3% (volume) in gasoline and 5%(mass) in diesel.
Moreover, in order to prevent the competition withfood supplies that has occurred recently, as well as
related to Post Kyoto Protocol efforts, an international
movement is occurring that seeks an appropriate
evaluation of CO2 reduction effects, including all
aspects of the process from the use of land to
manufacturing and transportation.19) In any event,biomass resources are not very abundant in Japan, so
their use must be limited. On the other hand, throughthe development of new fuel production methods,
there are expectations that we can contribute to its
use abroad.
If I am to make a few remarks on fuel-cell vehicles
that use hydrogen as fuel, the fact is that their CO2
reduction effect will not be great as long as the
manufacture of hydrogen depends on petroleum,
natural gas, coal and other fossil fuels. Eventually,systems for manufacturing and supplying hydrogenthat eliminate dependence on fossil fuels and achieve
great reductions in CO2 emissions must be realized. Ifthese conditions can be met, hydrogen could become
popular, but many other obstacles must also be
overcome, such as the storage and supply ofhydrogen, how it is carried in vehicles and its userfriendliness. Other obstacles are issues related to thefuel cell stacks themselves, including theirperformance, reliability and durability and cost
reduction. At present, long-term research anddevelopment must be continued with the support of
the national government. Specifically, 30 companies
have made a joint declaration targeting 2015 as thestarting point for the real expansion of thistechnology.20)

The sudden occurrence of the Great East Japan
Earthquake this past March 11, however, reduced theratio of nuclear power generation, which has been amajor source of power for EVs and PHEVs, andcaused an increase in the ratio of thermal power
generation. For this reason, although the trend is anincrease in the basic unit of CO2 emissions, thepotential exists for greater use of power generated
from renewable energy sources, including solar, wind,geothermal and biomass. In this respect, we can
expect that this will lead to eventual reduction in CO2emissions. In addition, many ideas have been
proposed for using the batteries of these vehicles tostore electricity generated by residential solar power
and as emergency power sources.

Daily lifestylerange

Power information
Weather information

Statistics/predictions/actual records

Power storage andconsumption plan

PHV・EV

PHV・EV
Home
Solar

EcoCuteheat pumpwater heater

Informationtransmission

Consumption pattern
prediction

Optimized energy
calculations

DB

Data collection

Home

Smart
phone

Auto & home

ToyotaSmart Center

PHV・EV

Storagebattery

Smart
phone

Fig. 6 Toyota Smart Center energy management
system—service will start in 2012 (Source: Toyota Motor
Corporation, October 2010)

4. Use of New Fuels and Energy Sources
The promotion of the use of new fuels and energy
sources to supplement the amount of gasoline anddiesel used is an important task in consideration ofsuch issues as reducing fuel consumption, diversifying
energy sources and taking measures against global
warming. As shown in Fig. 7, diverse options can besuggested to supplement fossil fuels. Electricity forelectric vehicles and hydrogen for fuel cells, as
discussed above, are also included in these.
As for renewable fuels, we have bioethanol andbiodiesel, which use biomass for their raw materials.17,18) These fuels can be utilized with minimal change tothe fuel infrastructure and to the vehicles themselves.
Moreover, since they are liquids, they are similar toand more compatible with conventional fuels.
Bioethanol is produced through the fermentation ofsugarcane, corn and other sugars and starches, aswell as cellulose-based raw materials using a variety
of processes. Biodiesel is produced from vegetable

Original resource

Well toDerivative

Fassil & neclear

Tank

Petroleum(oil sands and oil shale)

Renewable

Gasoline＊
Diesel＊
CNG, LNG

Coal

DME

Biomass
Waste products

Spark-ignition engine
vehicles (Hybrid)
Compression-ignitionengine vehicles (Hybrid)

MethanolFuel-cell vehicles(Hybrid)

HydrogenElectricity＊＊
Bioethanol

Solar power Hydropower
Wind power Geothermal

Applicable vehicle type

wheel

Natural gas

Nuclear power

Tank toElectric vehicles＊：Synthetics (GTL, CTL, BTL)＊＊： Including power generation

Biodieselby fossil fuelsFig. 7 Future paths of fuel and energy generation for vehicles-8-Prospects for Environmental and Energy-related Vehicle Technologies

In order to achieve weight reduction of vehicles in their
entirety, efforts must be made in a wide range fromstructural parts to the component level, including
making items more compact. Of course, these types
of weight reductions have the benefit of improving fueleconomy not only for passenger cars but also forcommercial vehicles.
As a specific example, I will introduce the World
Auto Steel project that has been undertaken since2008 by 16 major international steel manufacturers. In
this project for electric vehicles and hybrid vehicles,
they are seeking to reduce vehicle weights by about
30% to improve fuel economy by 20% while assuringsafety by using high tension and super high tension
steels that are 2 to 4 times as strong as previous
materials.21) These types of vehicles tend to become
heavy compared to gasoline vehicles, so reducing
their weight can be said to be necessary. Japanese
steelmakers have a lead in this field, and theapplication of their results is steadily beginning to
expand.
In addition, the use of aluminum and otherlightweight metals, as well as plastics, including
CFRP, is also being advanced. Hopes are high for theadvancement of the popularization of technologies
that use these materials as the benefits of each typeof material are utilized and the obstacles, which
include difficulties related to molding, finishing andjoining with different types of parts, increased costs
and the globalization of production, are overcome.
With these materials, another important issue will beassuring control of vehicle vibrations and noise as wellas collision compatibility between vehicles. Structural
optimizations that minimize harm to people and thedamage caused by heavier vehicles to lighter vehicles
during collisions are needed. I want to point out thatthese types of vehicle weight reductions also motivate
research and development for cutting-edge
technologies that seek to increase comfort, provide
collision safety and prevent accidents.

5. Techniques to Reduce Vehicle Weight
I want to emphasize that techniques that reduce
vehicle weight are also extremely important for fueleconomy improvements. Fig. 8 shows a comparisonof the energy consumption caused by vehicle weight,
rolling resistance and air drag while driving based on alight vehicle traveling city streets. Fig. 9 is an example
of a quantitative estimation that I prepared through thenondimensionalization of this data. It shows theimpacts that reducing these effects could have onenergy consumption when driving. These figures
make clear that, among these three factors, the effectof weight reduction on improving fuel economy could
be very large. It goes without saying that this is an
important technique for both hybrid vehicles andelectric vehicles. As shown in Fig. 10, weight
reduction not only improves the operation
performance of vehicles, but also enables thedownsizing of power systems and the reduction ofemissions purification loads, creating a virtuous circle.

Energy consumption share %

60
50
40

Driving mode: JC08
Energy consumption rate: 0.345 MJ/km
M：1,200kg
CdA：0.6m3
μr：0.01

28.4%

30

51.5%

27.5%
Proportion of energythat can be recaptured by regenerative
braking (equivalent
to energy reduction
ratio)

20.1%

20
10
0

[Rolling
resistance][Air drag][Acceleration
resistance](Weight reduction contributes to energy conservation)

Fig. 8 Energy consumption when driving a vehicle

1.4
Driving energy comparison

Driving mode: JC08
1.2

1.0

CdA
μ

Vehicle weight
reduction

M

0.8

0.6
0.6

0.8

1.0

1.2

<Body, steering & suspension>
Machining Production
Design Recycling Safety
Driving performance

1.4

Variables: M/M0, μ/μ0, CdA/CdA0

Power systemdownsizing andweight reduction<Engine>Combustion Supercharging
Exhaust purification
Transmission system Control<Electrification & hybridization>

Vehicle weight M0 = 1,200 kg
Rolling resistance coefficient μ0 = 0.01
Air drag coefficient×total projected area: CdA0 = 0.6

Reduction of environmental impacts
Energy conservation
Resource conservation
Safety assurance

Fig. 9 Resistance factors and required energy for driving
a vehicle

Fig. 10 Synergy between power system downsizing andvehicle weight reduction-9-

NTN TECHNICAL REVIEW No.79（2011）<Relative CO2 amount %>6. Future Prospects and Conclusion

0

20

40

60

80

100

■Current gasoline vehicle

After conforming to the planned final emissions
regulations in Japan, America and Europe by the
middle of the second decade of the 21st century,
developments and advancements will continue forgasoline and diesel vehicles with an emphasis on fueleconomy improvement technologies. I expect thatthey will continue to have an important role for at least
the next 20 years. In their advancement, assumingthat fuel characteristics improve, the combination andoptimization of elemental technologies related to
combustion technologies and exhaust after-treatment
technologies will be indispensable.
In addition to these conventional technologies, thereduction of vehicle weights and the use of hybrid andelectric vehicles as well as biofuels, for example, arenecessary. Fig. 11 shows the results of predictions
about future CO2 reductions resulting from thesetechnologies. In this chart, a ■ indicates a current
technology, a ● indicates a reduction resulting fromthe power system, and a ☆ indicates a reduction fromanother technology. The horizontal bars indicate therange of the effect depending on differences in thetechnology. This chart also shows us that the CO2
reduction potential of EVs is very high. At the Next
Generation Vehicle Strategic Research Meetingsponsored by the Ministry of Economy, Trade andIndustry, international standardization, establishingrecharging infrastructure and securing resources weresuggested as important strategies related to EVs andbatteries.22) The same research meeting provided aforecast of comparative sales shares of nextgeneration vehicles in 2020 and 2030, as shown in
Table 3. We can expect that it would be possible for
adoption to occur even more quickly with proactive
government support compared to if only businesses
make efforts.
Moreover, assuming the wider adoption ofIntelligent Transport Systems (ITS) that utilize
telecommunication technologies, which are expectedto advance even more in the future, reviewing the useof all types of vehicles and advancing performanceincreases are necessary. These include traffic flowsmoothing and controlling suitable traffic volumes,
increasing the efficiency of freight traffic, promoting
the use of public transportation, shifting to railtransportation, and reevaluating commercial habits
and lifestyle patterns that depend on automobiles(such as the promotion of green driving).
If we can advance these measurescomprehensively, the potential CO2 reduction ispredicted to possibly be about 50% by 2030 and 80%
by 2050, as shown in Fig. 12. To achieve these☆Future gasoline vehicle■Current diesel vehicle☆Future diesel vehicle☆Future gasoline HV☆Future diesel HV☆Future EV (light)）●Vehicle weight reduction●Biofuel use
Assumptions:• Total efficiency = fuel efficiency×vehicle efficiency• Calorie ratio of thermal power from fossil fuels in EV power sources: 60%• Vehicle weight reduction: 20–40%• Calorie-converted mixture ratio of biomass: 6–12%

Fig. 11 Comparison of vehicle CO2 emissions in the
2020s (Baseline: current gasoline vehicle)

Table 3 Projected sales shares of next generation
vehicles in 2020 and 2030
(Source: Next Generation Vehicle Strategic Research
Meeting, METI, March 2010)<Private sector effort model>
2020

2030

Conventional vehicles

80% or more

60〜70％

Next-generation vehicles

20% or less

30〜40％

Hybrid vehicles

10〜15％

20〜30％

EVs & plug-in hybrid vehicles

5〜10％

10〜20％

Fuel cell vehicles

Minimal1％

Clean diesel vehicles

Minimal〜5％<Government targets>
2020

2030

Conventional vehicles

50〜80％

30〜50％

Next-generation vehicles

20〜50％

50〜70％

Hybrid vehicles

20〜30％

30〜40％

EVs & plug-in hybrid vehicles

15〜20％

20〜30％

Fuel cell vehicles

〜1％

〜3％

Clean diesel vehicles

〜5％

5〜10％

Standard△30-40％ △45-55％ △65-85％△15-25△5-10△5-10△25-35△10-15
△5-10

100%
60-70

△35-45△15-252020<Non-fossil fuel energy use>
Electricity, biomass, CCS△10-2045-55
15-35

Present

Reduction method<Conventional vehicle fueleconomy improvement measures>
Increased power system efficiency
Hybridization, vehicle weight
reduction

2030<Vehicle use improvement andenhancement>
TCM, ITS, modal shifts,
Car culture changes

2050

Fig. 12 Projected medium and long-term reductions in
vehicle CO2 emissions-10-

Prospects for Environmental and Energy-related Vehicle Technologies

passenger cars, Ministry of Economy, Trade and Industry(http://www.meti.go.jp/report/downloadfiles/g50418b01j.pdf), April 2005.
8) Osaka, T., Daisho, Y., et al., Electric Automobile
Handbook, Maruzen, 2001.
9) State of California (USA) website(http://www.arb.ca.gov/msprog/levprog/levprog.htm).
10) US PNGV program(http://www.fueleconomy.gov/feg/pngv.shtml).
11) Kihara, R., Daisho, Y., Research on High-performance
Hybrid Vehicles, Sankaido, 2005.
12) Japan Hydrogen & Fuel Cell Demonstration Project(http://www.jhfc.jp/j/index.html), 2002–present.
13) Measurement methods for the exhaust gas and fueleconomy of plug-in hybrids, Ministry of Land,
Infrastructure, Transport and Tourism(www.mlit.go.jp/common/000046352.pdf), 2009.
14) Roadmap for Technological Development of Storage
Batteries for Next-generation Automotive Electric Power
Systems, NEDO
(https://app3.infoc.nedo.go.jp/informations/koubo/other/F
A/nedoothernews.2009-05-29.2374124845/), 2009.
15) EV & pHV Towns Concept, Ministry of Economy, Tradeand Industry(http://www.meti.go.jp/policy/automobile/evphv/index.html)
2009.
16) Ministry of Economy, Trade and Industry, Nextgeneration Energy and Social Systems Demonstration
Areas(http://www.meti.go.jp/policy/energy_environment/smart_community/community.html), April 2010.
17) Increasing the utilization of green transportation fuels,
Green Fuel Use Promotion Committee, Ministry of theEnvironment(http://www.env.go.jp/earth/ondanka/conf_ecofuel/), 2006.
18) Daisho, et al., The Frontline of Biodiesel (Revised
Edition), Kogyo Chosakai Publishing Co., Ltd., June 2008.
19) Report on the “Investigative committee on sustainability
standards and other matters concerning the adoption of
biofuel use,” Ministry of Economy, Trade and Industry(http://www.meti.go.jp/press/20100305002/20100305002.html), March 2010.
20) Toyota Motor Corporation(http://www2.toyota.co.jp/jp/news/11/01/nt11_0106.html),
Jan. 2011.
21) Efforts by WorldAutoSteel(http://www.worldautosteel.org/About.aspx).
22) Next-Generation Vehicle Plan 2010, Ministry of
Economy, Trade and Industry(http://www.meti.go.jp/press/20100412002/20100412002-3.pdf), March 2011.

predictions, medium and long-term outlooks forsecuring resources, conserving energy, and reducing
CO2 are necessary along with the execution ofnational support measures to realize them. While
keeping an eye on the international marketplace,including emerging nations, and energy policies, theindustry should also devise and execute technologicalstrategies accordingly.
NTN has been conducting the development andmanufacture of power transmission systems, related
bearings and similar parts for automobiles. Recently,
however, the company has also been advancing thedevelopment of technologies related to the trend
toward vehicle electrification. These are indispensable
technologies for the next generation of vehicles, and I
have great expectations for your future advancementswith them.

Reference
1) Forms of vehicle exhaust gas reduction measures in the
future, reports 2–10 of the Central Environmental Council,
1997–2010.
2) World Energy Outlook 2010, IEA(http://www.iea.org/weo/2010.asp), 2010.
3) Fiscal 2010 fuel economy standards for passenger carsand other vehicles, Ministry of Economy, Trade andIndustry, and Ministry of Land, Infrastructure andTransport(http://www.meti.go.jp/feedback/data/iscar00j.html), 2010.
4) Final arrangement of fiscal 2015 fuel economy standards
for cars and other vehicles, Ministry of Economy, Tradeand Industry, and Ministry of Land, Infrastructure andTransport(http://www.mlit.go.jp/kisha/kisha07/09/090202_2_.html).
5) Fiscal 2010 fuel economy standards for passenger carsand other vehicles, Ministry of Economy, Trade andIndustry, and Ministry of Land, Infrastructure andTransport(http://www.meti.go.jp/feedback/data/iscar00j.html), 2010.
6) Daisho, Y., Environmental and regulatory trends related toautomobiles—the past ten years, Journal of the Society ofAutomotive Engineers of Japan, Jan. 2010.
7) Report of an investigative commission related to the
popularization of and future prospects for clean diesel<Author biography>[Research fields]• Engine combustion, exhaust purification, efficiency improvement andnew fuel utilization technologies• Manufacture, performance evaluation and other studies of electric,
hybrid and fuel-cell vehicles

Yasuhiro Daisho
Professor, Faculty of Science and Engineering, Waseda University
Dean, Graduate School of Environment and Energy Engineering
Director, Environmental Research Institute
Professor, Department of Modern Mechanical Engineering, Graduate
School of Creative Science and Engineering
1976
1980
1985

[Academic society and committee affiliations]• Vice President, International Federation of Automotive Engineering
Societies (FISITA)• Advisor, Ministry of the Environment Central Environment Council• Member, Ministry of Land, Infrastructure and Transport
Transportation Policy Council and Social Capital Policy Council• Member, Advisory Committee on Energy and Natural Resources• Director, Japan Automobile Research Institute• Member or chairman, other committees related to vehicles and theenvironment and energy, etc.

Received Ph.D. from Faculty of Science and Engineering,
Waseda University Graduate School
Became Assistant Professor of Department of Science andEngineering, Waseda University
Became Professor of Department of Science and Engineering,
Waseda University-11-

NTN TECHNICAL REVIEW No.79（2011）[ Perspective ]Market and Technology Trends in the Automotive Industry

Makoto OKASAKA*

After the collapse of Lehman Brothers, the Ministry of Economy, Tradeand Industry announced a new strategy to build long-term and short-term
responses for the automotive industry and society as a whole in April
2010. The results of the Next-Generation Vehicle Strategy Research
Association were published as the Next-Generation Vehicle Plan 2010.
We have begun to see signs of revival even in the domestic auto industry,and discrepancies with this scenario have begun to arise in 2011.
In this report, we look at the present and future market and technology
trends in the automotive industry as a developer of automotive bearing
sand related products.

has changed from the export of vehicles produced
domestically to the production of both parts andvehicle bodies in the emerging markets themselves.
With these changes, however, the automobile market
is continuing to expand.
At the same time, out of concern for preventing
global warming and other worldwide environmental
problems, as well as in response to energy concerns
due to the steep rise in crude oil prices, thedevelopment of electric vehicles (EV), hybrid vehicles(HEV) and fuel cell vehicles (FCV) has begun toflourish in many countries. On the other hand, in
response to changes in the automotive market,
development of technologies for conventional vehicles
with gasoline and diesel engines that consume fossil
fuels has also become vigorous. These include ways
to increase fuel economy, reduce weight and purify
exhaust gases, as well as to promote greater
electrification and control-by-wire utilization.

1. Introduction
In April 2010, the Ministry of Economy, Trade andIndustry (METI) officially published the Next
Generation Auto Strategy 2010.1)
This strategy is the product from the Research
Conference on Next Generation Auto Strategy wereexperts from METI and the Ministry of Land,
Infrastructure and Transport (MLIT), automakers andautomotive component manufactures in Japan took
part, in order to develop a new strategy for medium andlong term activities to be taken by the entire society inJapan including Japan's automakers and automotive
component manufactures that were experiencing
difficulties after the Lehman Brothers shock.
In the global automotive industry, a period of great
success with six consecutive years of record-breaking
car production numbers continued until the Lehman
Brothers shock, and then, a rapid decline in demandfollowed, and both the US Big Three and the Japanese
automakers fell into the red. Later, numerous aid
measures and economic stimulus policies havestarted the recovery of demand, but it is still stagnant
at a low level.
In contrast, the emerging markets, in particular,
Brazil, Russia, India and China (the BRIC countries),
experienced temporary stagnation, but theirrecoveries were remarkable. In addition, as a result of
the changing conditions of the global economy withthe acceptance of a high yen, the Japanese industry

2. Changes in Automotive Market[Bipolarization of the automobile market]
Within the global automobile market, the Chinese
market has grown and surpassed the North American
market, which includes Canada and the United States,
exceeding 13,790,000 vehicles in 2009. Meanwhile,
the Japanese domestic market was stuck at 8,000,000
vehicles.
In the Next Generation Auto Strategy 2010, trends* Automotive Business HQ. Automotive Engineering Dept.

-12-

Market and Technology Trends in the Automotive Industry(Vehicles in tens of thousands)

67,960,000 vehicles (2008)

8,000
7,000

[2] Emerging markets

6,000
5,000

Emerging
markets

31,000,000
vehicles

4,000

Japan

5,080,000
vehiclesEurope15,600,000
vehicles

3,000
2,000

1,000

North America

0
1930

1960

1972

1976

1980

1984

1988

1992

1996

2000

2004

Increased population thatbuys cars as economies
grow[1] Industrialized nation
markets
Mature stable markets withdiversifying user andsociety needs

16,200,000
vehicles2008 (year)

Fig. 1 Changes in the global auto market Reference: Next Generation Auto Strategy 20101)

country market. Initially, carmakers did not foresee theseverity of these conditions.
Furthermore, in addition to countermeasuresagainst salt damage from snow-melting agents andmuddy water, since the frequency of car-washing toclean off mud that adheres to vehicles is also great,consideration for high-pressure car-washing is also
necessary.

in the global automobile market are summarized asshown in Fig. 1. The growth of the markets in India,
China and other emerging countries in recent years isparticularly notable.
The rapid growth of the emerging markets, looked
at in another way, suggests the beginning of abipolarization of automobile markets into those of theindustrialized countries and those of the emerging
countries.

[Changes in power sources: breaking awayfrom fossil fuels][Responding to markets with difficult
environmental conditions]

Based on the Copenhagen Accord from theprevious 15th Conference of the Parties (COP15) of
the United Nations Framework Convention on Climate
Change, Japan has proposed a target of reducing
greenhouse gases by 25% from 1990 levels by 2020.
In addition, the price level of fossil fuels for vehicles
has remained high with the increased cost of crude oilsince the 2008 Lehman Shock, as well as because of
the expansion of thermal power generation due toproblems related to nuclear power generation in
Japan.
As a result, development of alternatives, such as
electric vehicles (EV), hybrid vehicles (HEV) and fuelcell vehicles (FCV), has become active in various
nations. At the same time, various carmakers havealso begun developing various means of improving
fuel efficiency for the internal combustion engines thatwill continue to use fossil fuels.
Both government and private parties are predicting
the popularization of these types of next-generation
vehicles. The Ministry of Economy, Trade andIndustry's Conference on Next Generation Auto
Strategy predicts that, with private efforts, the usage

In the markets of emerging nations where demandis increasing rapidly, in addition to responding toenvironments that range from extremely cold
temperatures to very hot and humid conditions,
measures that are sufficient to deal with mud arenecessary because there are many regions that haveinsufficient road infrastructure.
As an example, Fig. 2 illustrates road conditions ina region with a difficult environment. The deep ruts
and mud on the road surface create a difficult
environment that is hard to imagine in an industrialized

Fig. 2 Poor road conditions in area with difficult environment-13-

NTN TECHNICAL REVIEW No.79（2011）

ratios shown in Table 1 will be achieved in 2020 and
2030 for these vehicle types.
In addition, the Conference on Next Generation
Auto Strategy indicates that groupings can be madefor these next-generation vehicles according to vehicle
size and driving distance capability, as shown in Fig. 3.
In summary, they can be grouped into EVs as
commuters for transporting small numbers of people
over short distances within a community, HEVs fordriving medium distances, and FCVs for traveling
relatively long distances.

<European market>
18,000,000
16,000,000
14,000,000
12,000,000
10,000,000

Units/year 8,000,000
6,000,000

4,500,000 4,000,000
4,000,000

2,000,000
03,500,000
3,000,000
2,500,000
2,000,000
1,500,000
1,000,000
500,000[Changes in transmission types: reducers vs.
transmissions more gear stages]

0
'00

'01

'02

'03

'04

'05

'06

'07

'08

'09

'10

'11 '12

'04

'05

'06

'07

'08

'09

'10

'11 '12

'06

'07

'08

'09

'10

'11 '12<Japanese market>

If the power source is an EV motor with a wide
torque band, as used in EVs and FCVs, atransmission is unnecessary and a reducer is utilized.
For HEVs and conventional fossil-fuel automobiles
that use internal combustion engines with narrow
torque bands, however, multistage transmissions orcontinuously variable transmissions are necessary.
In addition to manual transmissions (MT), automatictransmissions (AT) and continuously variabletransmissions (CVT), other transmission types include
automated manual transmissions (AMT) and dual
clutch transmissions (DCT).
As an example of the market trends for these typesof transmissions, Fig. 4 shows the changes intransmission type in front-engine, front-drive (FF)

Units/year
9,000,000
7,000,000
8,000,000
6,000,000
5,000,000
4,000,000
3,000,000
2,000,000
1,000,000
0
'00

'01

'02

'03

<North American market>
Units/year
10,000,000
9,000,000
8,000,000
7,000,000
6,000,000

Table 1 2020 & 2030 forecast for car types1)

5,000,000

2020

2030

Conventional vehicles

80% or more

60〜70%

Next-generation vehicle

20% or less

30〜40%

Hybrid vehicles

10〜15%

20〜30%

Electric vehicles
Plug-in hybrid vehicles

5〜10%

10〜20%

Fuel cell vehicles

Minimal1%

Clean diesel vehicles

Minimal〜5%4,000,000
3,000,000
2,000,000
1,000,000
0
'00

'02

'03

'04

'05

Fig. 4 Changes in transmission type used in FF cars 2)

Clean diesel vehicle domain

Large

HV & PHV domain
Vehicle size'01

EV domain(At the stage of market acceptance(EV) or introduction PHV))FCV domain(Research anddevelopment stage)(At the stage ofmarket introduction)
Short distance,
local commuters

Driving distance
Long

Fuel

Electricity

Gasoline, diesel, CNG, LPG,
biofuel, synthetic fuel, etc.

Fig. 3 Future groupings by vehicle type 1)-14-

HydrogenMarket and Technology Trends in the Automotive Industry

than they have. In contrast, the market for passenger
cars is expected to continue growing globally,
primarily in emerging markets, with the number ofpassenger cars alone exceeding 100 million by 2030.

passenger cars in the markets of Europe, Japan andNorth America. Since lowering the engine speed
should improve fuel economy, multistagetransmissions with greater numbers of transmission
gear stages have increased rapidly in recent years.
In the European market, MTs continue to be
popular, but AMTs and DCTs, which make driving
easier and have good acceleration and fuel economy
characteristics, are rapidly increasing in popularity.
In the Japanese market, CVTs are steadily gaining
popularity and ATs are shifting to 5- and 6-speed
multistage versions.
On the other hand, in the North American market,
ATs are as dominant as ever, but the common
number of stages in them has increased from 3〜5
speeds to 5〜6 speeds.

[Relative loss of interest in cars among young
people]The ratio of licensed drivers among men andwomen between 18 and 24 years old has not changed
much from 30 years ago in Japan. JAMA, however,has conducted a survey related to the "loss of interestin cars among youth." 4)
The results of this survey, which ranked
commodities according to the level of interest ofcollege-aged respondents of that generation, showedthat younger generations had decreasing interest in
cars.
･ College students 20 to 30 years ago(40〜50 years old at present): 7th･ College students 10 to 20 years ago(30〜40 years old at present): 10th･ Current college students (18〜24 at present): 17th

[Changes in growing passenger car markets]
While the bipolarization of the markets ofindustrialized nations and emerging nations and thetransition to EVs, HEVs, FCVs and other nextgeneration vehicles are occurring, the recovery ofgrowth in the passenger car market, excluding trucks
and buses, since the Lehman Shock has been fueledby India, China and other emerging markets.
According to the totals from the Japan Automobile
Manufacturers Association (JAMA), in fiscal 2010,global production of automobiles, including trucks andbuses, was 77,610,000 vehicles, of which 58,270,000
were passenger cars. Given the ever-increasing
production of passenger cars in India, China and othercountries, one private research institution has
predicted that the number of passenger cars alone will
exceed 72 million in 2020, as shown in Fig. 5.3)
In summary, the prediction is that the markets forpassenger cars in Europe, Japan, North America andother industrialized nations will not grow any more

Other countries
Japan

India
EU

Among the recent younger generation, which does
not have much interest in cars and has low buying
power, cars are merely one means of transportation,and they tend to prioritize economic issues (good fueleconomy, low vehicle price and inexpensive
maintenance). From the results of this survey, we cansee that the trend will be for inexpensive compact cars
to sell well.

3. Recent Engineering Trends in
Automotive Industry
From these market trends, we can see that for theJapanese automotive industry to remain competitiveand survive as a top global industry, developing thefollowing technologies and advancing well-balanced
production strategies in emerging markets are crucial.
･ Development of EVs, HEVs, FCVs and other nextgeneration vehicles that combine energy
conservation and environmental protection･ Development of fuel economy improvements andweight reduction for internal combustion engine
vehicles to maintain competitiveness in this field
Table 2 shows engineering trends for improving
electrical economy and fuel economy in theautomotive industry.

China
United States(Millions of vehicles)
80

72,000,000

70

54,000,000
60
50
40
30
Low growth

20
10

2020
(Year)

2018

2016

2014

2012

2010

2008

2006

2004

2002

2000

1998

1996

1994

1992

Low growth

1990

0

Fig. 5 World car sales forecast 3)-15-

NTN TECHNICAL REVIEW No.79（2011）

Table 2 Recent trends in the automotive industry
Area of application

Engineering trend¡Low friction motors¡High-RPM compact motors ¡Motor bearing torque reduction

Fuel-related
improvements¡High-pressure fuel injection¡Lean burn¡Miller cycle¡Hybrids

Precision
control¡Stop idling¡Cutting fuel supply during deceleration

EV motors¡Incorporation of rolling bearing in fuel pump

¡Fuel-air ratio control and ignition control¡Electric drive of auxiliary equipment¡Downsizing by introduction of turbo superchargers¡Variable pressure mechanisms
Application ofEngines
new mechanisms ¡Variable valve movement mechanisms (variable valve timing and variable lift) ¡Variable cylinders (cylinder rest)¡Low viscosity engine oil
Friction reduction

Power
trainsAT improvements

Chassis Weight reduction
Others¡Dry sump lubrication¡Incorporation of rolling bearing onto crankshafts¡Oil pump size reduction¡Incorporation of rolling bearing onto camshafts¡Low-tension auxiliaries drive belt systems¡Incorporation of rolling bearing onto balancer shafts¡Reduced friction loss in valve powertrains¡Incorporation of rolling bearing onto chain levers¡CVTs with secondary transmission ¡DCTs¡ATs with more stages¡Ball-screw driven automatic transmissions (AMT)¡ATs with built-in EV motors¡Increased use of resins and high-tension steels¡Reduction in hub bearing weight¡Low μ tires¡Zero lift bodies¡Reduction in hub bearing torque¡Reduction of air drag coefficient¡Reduction of glass area

¡Control-by-wire[Low torque bearings for EV motors]

4. NTN's Efforts for Next-Generation
VehiclesAs for the bearings of next-generation vehicles
using EV motors, including EVs, HEVs and FCVs,
NTN has developed two types of low torque bearings.

[In-wheel motor drive systems]
Among electric vehicles (EV), which are nextgeneration vehicles, the type with the best
performance is the EV with in-wheel motors that carry
all of its drive parts inside its axle.
As shown in Fig. 6, in-wheel motors require
coordinated control between the left and right wheels,
but they also have the latent potential to create newautomobile driving features because each wheel canbe controlled separately. For this reason, NTN has
been developing its propriety in-wheel motor drive
systems that include controls using, for example,
inverters.

<Grease-lubricated type>Fig. 8 shows a grease-lubricated type of low-torque,sealed, deep-groove ball bearing for EV motors.
This bearing uses specially-designed cage pockets
to scrape off excess grease and prevent the influx ofexcess grease onto the rolling contact surface,
reducing stirring resistance. Furthermore, by using
grease that has a base oil and thickener with low
viscosity and excellent heat resistance properties,
rotation torque is reduced 50% compared to previous
products.

[One-motor-powered EV systems]<Oil-lubricated type>

NTN is also developing one-motor-powered EV
systems that can readily replace internal combustion
engines because the structures of car bodies for thistype of EV system are the simplest among electricvehicle forms. Fig. 7 shows an outline of this type.
This one-motor-powered EV system was developed
through a combination of an automatic 2-stage
reducer and constant velocity joint that fully reflectsconsiderations related to an EV motors output
characteristics. This system is capable of morecommon use as it can be used not only in front-wheel
drive electric vehicles, but also for the rear wheels onelectric four-wheel drive vehicles.

Fig. 9 shows NTN's oil-lubricated, high-speed lowtorque deep-groove ball bearing for EV and HEV
motors.
In the future, oil-cooling systems will be used to cool
EV motors in order to make them more compact byenabling super-high-speed rotation.
As a result, motor shaft support bearings will be oillubricated, so we developed this bearing to meet thisneed.
Stirring resistance inside the bearing is reduced by
optimizing the form of the cage and controlling theinflux of oil into the bearing. Furthermore, cage
deformation caused by centrifugal force during superhigh-speed rotation can be inhibited by using a resin--16-Market and Technology Trends in the Automotive Industry

Motor

Hub

Ti
re

H/B

IWM
unit

d
ee
sp

Accelerator
pedal

Reducer

AccelerationcommandDC
power
In-wheel motor system
information

Brake
pedal

Steer
commandInverter

VCU

DecelerationcommandMotor rotor angle
Motor rotor angle

Left and right wheel
torque commands

DC
power
AC power
H/B

Battery

IWM
unit

Tire speed

In-wheel motor
systemFig. 6 In-wheel motor drive system

EV

Drive motor
Transmission

Fig. 7 One-motor-powered EV system

Oil-lubricated type

Grease-lubricated type

Duplex resin-made snap cageLow torque grease
Grease not readily scraped away

Torque reduction
Oil influx into the bearing
is controlled and stirring
resistance is reduced

Grease readily scraped off

Torque reduction
Conventional
cage

Grease not readily scraped away

Centrifugal force
countermeasure

Steel balls are held by the outercircumferential of pockets.

Low torque Non-contact
cage
seal

Optimized design
reduces rolling viscosity
resistanceCage deformation caused
by centrifugal force iscontrolled by using a
combined structure

NTN’s newly
developed cage

Fig. 8 Low-torque sealed deep-groove ball bearing
for EV motors

Fig. 9 High-speed low-torque deep-groove ball bearings

made duplex cage for higher bearing rigidity; thereby
the bearing is capable of positively withstanding
super-high-speed applications at 30,000 min-1 andreduces rotary torque by more than 40% compared toconventional bearings.

-17-

NTN TECHNICAL REVIEW No.79（2011）attention as items that improve fuel economy
because, compared to sliding bearings, they haveconsiderably lower torque at lower speed ranges. This
reduces engine friction loss when a vehicle starts tomove and when it accelerates, which are times whenfuel consumption is high.

5. NTN's Efforts for High Fuel Economy
VehiclesThe main impediments to better fuel economy forvehicles, taking automatic transmission vehicles as an
example, are shown in Fig. 10.
In order to increase fuel economy in hybrid vehicles
and vehicles that use the fossil fuels of gasoline anddiesel, NTN is developing optimized low-torquebearings for each place where they are used, as well
as developing modular products for vehicle
electrification.

Enginefriction27%

Air drag
19%[Items that improve the fuel economy of fossil
fuel engines]
Fig. 11 shows examples of items that NTN has
developed for the purpose of improving the fueleconomy of fossil fuel engines that use gasoline anddiesel.
The rolling bearings introduced here are receiving

Needle roller bearing
for rocker arms

Torque converter
4%

A/T friction
9%

Vehicle weight20%

Tire rolling
21%

Fig. 10 Impediments to fuel efficiency (in HWY mode) 5)

Shell needle roller bearing
for camshaft

Ballscrew for continuously
variable valve mechanism

Continuously variable
valve mechanism

Roller lifter
for fuel injection pump

Segmented needle bearing
for crankshaft

Low-tension chain tensioner

Fig. 11 NTN's products for improving the fuel economy of combustion engines 6)-18-

Market and Technology Trends in the Automotive Industry

motorcycles, which frequently start travel andaccelerate, fuel economy improvements of 1-3% havebeen achieved in actual motorcycles.
Moreover, the energy conservation effect of using
several of these low-torque ball bearings in the EV
motor and reducer on electric motor cycles has been
confirmed to be a 35W electricity consumption
reduction, which is equivalent to the amount used by
the front light for illumination.
Furthermore, in place of the conventional tapered
roller bearings for the differential side, we have a
deep-groove ball bearing for high thrust loads. This isan altered form of a deep-groove ball bearing that has
a thrust load capacity equivalent to that of lowgradient tapered roller bearing and reduces torque by
56% compared to tapered roller bearings, as shown in
Fig. 14.

In addition, as a fuel economy improvement item forthe engine valve system, a low-friction chain lever, as
shown in Fig. 12 is adopted; where each raceway
surface on the chain lever for applying tension to thetiming chain as well as on the chain guide for guiding
the timing guide is altered from the conventionalsliding type to rolling type.
Previously, the chain lever and chain guide, whichputs tensile load on the timing chain and guides it, hastypically been made of cast iron. Recently, the use oflightweight alloys and polyamide resin materials has
begun, but the drag torque acting on the chain is still
great because the contact mode between the chain
lever and chain guide with the chain is sliding contact.
NTN focused on this issue and has increased
rotating contact with the chain and reduced friction
torque by up to 18% at low speed ranges by placing a
lot of thick-plate shell needle bearings in multiple
locations on each of the chain levers and guides. As aresult of this, a fuel economy improvement effect of alittle less than 1% has been confirmed by measuring
fuel economy in actual engines.
Shell bearing

Non-sealed
low-torquedeep-grooveball bearing

Fig. 13 NTN's bearing product for transmission forimproved fuel-efficiency
Chain lever

Chain guideFig. 12 Low-friction chain lever

18

21.1

Newly developed product

φ40

φ80

φ80

NTN has been at the vanguard in the developmentof low-torque tapered roller bearings for many years.
However, even though further torque reduction had
been thought to be difficult, we have beaten thecompetition in developing a low-torque deep-grooveball bearing as an item to reduce transmission friction
loss further.
As an example of this, as shown in Fig. 13, we havedeveloped a non-sealed low-torque deep-groove ballbearing. The insides of the cage pockets of deepgroove ball bearings are made concave, reducing thefriction loss at the contact surface between the balland the cage by 25%.
In addition to the application of this bearing in
automotive transmissions, use of them in all the axial
bearings for the cranks, cams and transmissions of

φ40[Items that improve the fuel economy ofpowertrains]

Tapered roller bearing 30208

Rotation torque N・m

0.25

0.20

0.15

56% reduction

0.10

0.05
0

Newly developed
product

Tapered roller bearing
30208

Fig. 14 NTN's deep-groove ball bearing for high thrust loads-19-

NTN TECHNICAL REVIEW No.79（2011）

This deep-groove ball bearing capable of high thrustloads, as a differential bearing with reduced torque, is
expected to contribute to improving the fuel economy
of vehicles.

6. Efforts in Response to Electrification
In EVs, FCVs and other next-generation vehicles,internal combustion engines are no longer used as
power sources for energizing the auxiliaries;
consequently, the negative pressure from an engine
when drawing in air can no longer be used.
Moreover, for engines with high fuel economy and
HEVs that continue to use fossil fuels; auxiliary
equipment, including hydraulically operated variable
valve mechanisms and power steering, aresuperseded with electric mechanisms and the oilpumps that can inhibit fuel economy improvements
are replaced with smaller capacity pumps whose
capacity is just sufficient to lubricate the internal
combustion engine.
Furthermore, hydraulic foot brakes and handparking brakes that were boosted by engine negative
pressure are also being replaced with electricmechanisms, and automatic transmissions that use
electric shifting mechanisms have also been adopted
by a number of vehicles.
We predict that the electrification of these types ofauxiliary equipment and control systems will develop
as they are and be transformed into control-by-wireequipment. Taking these conditions into account, NTN
is anticipating the electrification and future control-bywire transitions in next-generation vehicles and isdeveloping a variety of modular products in responseto electrification needs.

[Items that improve chassis and affect the fuel economy]
Representative examples of items that improve thechassis for better fuel economy, which primarily
reduce hub bearing friction and weight, are shown in
Fig. 15.
Recently, hub bearings are beginning to be adopted
even on vehicles in regions with poor driving
conditions. Certain bearings that incorporate seals
with a dual side lip structure have been used, toimprove performance in severe mud-water conditions.
By changing the seal to a rubber material and revising
the form, however, not only is mud-water resistance
improved, torque is also reduced, contributing to
improved fuel economy for the vehicle.

Rolling friction reduced●Use of composite corner radius

Seal friction reduced●Use of new rubber material●Optimization of lip form

Rolling friction reduced●Reduced preload range

Seal friction reduction
Conventional
designSide lipNew
designThe vehicle market has overcome numerous trials
in the past and will continue to grow even as thelocations and forms of production change. The mostimportant things in the automotive industry are being
first to predict the latest market needs of each era andplanning and developing novel new products.
Immediately, based on a new theme of improving
electric economy and fuel economy, developments
are advancing with efforts toward production in
emerging markets and efforts toward electrification in
the form of EVs and other next-generation vehicles, aswell as efforts for technologies that improve the fueleconomy of fossil fuel vehicles, which are thefoundations for expansion of the automobile market.
After these have advanced, around the time whennext-generation vehicles have become mainstream, I
believe that the future development of vehicles will
once again begin, focusing on their original purposesof reaching destinations quickly, safely, in comfort andat will. At NTN, we intend to continue developing new
products for automobiles in the future.

New surface
shape

Slinger
Radial lip

7. Conclusion
Side lips
Double lip
design●Use of new rubber material

Fig. 15 NTN's low friction hub bearing-20-

Market and Technology Trends in the Automotive Industry

Reference
1) Ministry of Economy, Trade and Industry, Next
Generation Auto Strategy 2010, April 12, 2010.
2) Sanei Shobo, Motor Fan illustrated, Vol. 21, June 15,
2008. (Data source: CSM Worldwide/Jatco CTI
Symposium, June 12-14, 2007, Detroit, MI, USA.)3) Kitagawa, F., The automotive industry after 2010--what
we can see next, Nomura Research Institute, Ltd.,
Feb. 5, 2009; "Fuel economy strategies and newly
developing country strategies: the Run-up to 2025"
and "HIS Automotive's automobile market forecast,"
Automotive Technology, Nikkei Business Publications,Inc., July, 2011.
4) Japan Automobile Manufacturers Association, Inc.,"Passenger car market trend investigation--attitudes
toward cars of generations entering the car market,"
March 2009.
5) Industry and Science Systems Co., Ltd., Engine
friction reduction technology text, Engine Friction
Reduction Technology, Fujita, Y. (Nissan Motor
Company, Powertrain Division Headquarters); Ryouno,
Y., NTN Technical Review, Vol. 76, P. 80-87, 2008.

Photo of author

Makoto OKASAKA
Automotive Business HQ
Automotive Engineering Dept.

-21-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Paper ]In-Wheel Motor System

Yuichi ITOH*
Kayo SAKAI*
Yusuke MAKINO**
With increasing global environmental concerns, much attention has been paid toelectric vehicles for their high environmental performance. In-wheel motor drive
systems for electric vehicles have the advantages of effectively maximizing interior
space and providing vehicle driving performance due to the independent traction control
ability of each wheel. NTN has developed an in-wheel motor system with a sensorbased control system for the dawning electric vehicle age. This paper describes
evaluations of its decelerator and motor. In addition, it reports on results of vehicle tests
with the developed system.

independently but also of directly transmitting thedriving force to the tire without the intervention of adrive shaft. This makes it possible to achieve highlyresponsive control of the driving force. Making themost of this drive characteristic allows the vehicle
driving performance to be improved significantly.
Noting that IWM systems provide driving systems
suitable for next-generation EVs in this way, we havedeveloped the NTN-IWM System that incorporates an
IWM, which is composed of a reducer and a motor,
and a control system linked with sensor-based
information.
This paper reports the results of evaluation tests of
the IWM reducer and motor sections, and of a roadevaluation test done on an IWM-equipped vehicle.

1. Introduction
Under the current circumstances in which more
attention is being paid to environmental performanceand problems concerning energy, the automobile
industry, academic research institutes and a numberof similar organizations are engaged in thetechnological development of electric vehicles (EV) asa next-generation automobile candidate.
Two different drive systems are available for EVs.
The first is the one-motor system in which a motor isinstalled in the body and the power is transmitted toboth the left and right wheels via a differential device
and a drive shaft. The other is the in-wheel motor(IWM) system in which motors are installed in thewheels.
Since the motors and reducers are installed directlyin the wheels, an IWM system does not need adifferential, drive shaft or other parts1–8) when compared
to a one-motor system. Free of the layout restrictions
caused by these drive system components, the degreeof freedom in vehicle design is increased greatly by an
IWM system. In addition, with the removal of thedifferential and the drive shaft, energy loss is reduced
and this allows for a reduction in electricity
consumption required when running, thereby allowing
the driving range per charge to be increased.
Furthermore, this system is capable of not only
controlling the driving force of individual wheels

2. Specifications and structure of thedeveloped product2.1 Specifications
With a compact automobile with a displacement of1500 cc in mind, we designed an IWM system
assuming installation in the two rear wheels of suchan automobile. We adopted a wheel size of 15 inches.
The appearance and specifications of the IWM
developed are shown in Fig. 1 and Table 1.

**EV System Division Driveline System Engineering Dept.
**EV System Division Control System Engineering Dept.

-22-

In-Wheel Motor System

Cycloid reducer

Motor housing
Cooling fins

Reducer housing

Fig. 1 Appearance of in-wheel motor
Hub

Table 1 Specifications of in-wheel motor
Maximum output

30 kW

Maximum torque

500 NmMaximumrotational speed

1350 min-1(equivalent to 150 km/h with a 0.6m tire diameter )Reducer type

Cycloid reducer system

Reduction ratio

1/11

Motor typeRadial gap type permanent magnet synchronous motor

Cooling system

Air-cooled

Output shaft
Oil pump

Reducer input shaft

Hub section

Rotor shaft
Rotor
StatorReducer section

Motor section

Fig. 2 Cross-section of IWM

stator (including the coil) and rotor. Furthermore, tohelp actively cool the stator and lubricating oil, thehousing has cooling fins on its outer periphery andside faces so that the IWM is air-cooled.
Automatic transmission oil has been adopted as the
lubricating oil to reduce the pipeline drag with the oilbeing pressure-fed by the built-in oil pump, and to
reduce stirring loss due to rotating members in therotor and reducer, as well as to ensure sufficientdurability for the bearings and reducer.
In developing the IWM, we repeated a variety of
durability tests such as water-submersion tests andvibration tests under severe conditions, which wereidentical to those under which NTN hub bearings aretested, to ensure sufficient reliability during actual
running on roads.

2.2. Structure
Fig. 2 is a cross-section of the IWM system, which
has a structure that aligns the hub, reducer section
and the motor section in a series configuration.
Compared with other driving systems, an IWM systemequipped with the driving section in the wheels haslarger unsprung mass, putting it at a disadvantage interms of vehicle stability and riding comfort. For thisreason, reduced physical size and weight is an
important challenge in developing an IWM system.
In an IWM system, the motor section accounts for asignificant portion of its weight, and since the physicalsize of the motor is usually dependent on the
maximum torque it has to develop, reducing therequired motor torque is effective in reducing theweight of the IWM system. During development
reduction of the physical size and weight of the IWM
system was achieved by adopting a reducer to reduce
the torque required of the motor.
The lubrication system is based on an internal
circulation scheme in which a built-in oil pump driven
by the output shaft force-feeds the lubricating oil. The
force-fed oil passes through an oil passage installedinside the motor housing and flows through an internal
oil passage installed inside the rotor shaft and the
reducer input shaft that rotates in conjunction with therotor shaft and reaches the motor and interior of the
reducer. This forced oil circulation not only lubricatesthe bearings and reducer but also cools the motor

2.2.1 Reducer section
A cycloid reducer equipped with a K-H-V-type
planetary reducer mechanism 9) has been adopted that
not only requires less space but also provides a highreduction ratio. To reduce loss within the reducer,
roller bearings have been adopted in the internal gearthat contacts the oscillating-rotating external gear andthe output shaft. 7), 8)

2.2.2 Motor section
Table 2 shows the specifications of the motor. A
concentrated winding system has been adopted toshorten the axial length. Using both magnetic field
analysis and stress analysis, we derived thespecifications for the shapes of the stator and rotor-23-

NTN TECHNICAL REVIEW No.79（2011）cores and for the layout of permanent magnets inorder to achieve a reduction in the power loss in the
motor section to help realize an air cooled system aswell as high speed rotary motion.
Table 3 shows the specifications of the inverter.
The inverter is provided with higher reliability throughenhancements of a self-diagnostic function and circuit
checking.

Cycloid reducer
Input shaft bearing
Oil seal

Table 2 Motor specifications
Motor typeRadial gap type permanent magnet synchronous motor

Winding system

Concentrated winding

Maximum torque45 Nm

Maximum output

30 kW

Maximum rotational speed

Hub
Output shaft

Input shaft to Input shaft
the reducer

Fig. 3 Cross-section of test reducer (including the hub)-1

15,000 min

Table 4 Test conditions for reducer efficiency measurement
Table 3 Inverter specifications

Maximum input shaft rotational speed

15,000 min-1

DC 400 V

Maximum input shaft torque

45 Nm

Control power supply voltage

DC 12 V

LubricantAutomatic transmission oil

Output30 kW×2

Lubricant supply temperature

70˚C

Carrier frequency

10 kW×2

Lubrication system

Forced oil supply from the outside

Drive system

Sinusoidal wave PWM

Cooling system

Water-cooled

Drive power supply voltage

running of the IWM, was taken as the reverse input
condition.

3.1.2 Test results
Figs. 4 and 5 show the results of the measurementof the reducer transmission efficiency under theforward and backward input conditions. These chartsshow the efficiency of the reducer alone, with thepower loss from the hub, the bearing of the input
shaft, and oil seal in contact with the input shaft
excluded. The horizontal and vertical axes denote,
respectively, the rotational speed and the torque of the
input shaft. The sign of the input shaft torque ispositive for the forward direction input condition andnegative for the reverse input condition.
The maximum efficiency was 98% under theforward input condition and 97.5% under the reverse
input condition. With no noticeable difference
observed in the transmission efficiency between theforward direction and the reverse direction, about thesame efficiency was obtained for both power runningand regenerative running.

3. Bench Test (Efficiency Measurement)
3.1 Reducer section
3.1.1 Sample and test conditions
Fig. 3 shows the structure of the sample used inmeasuring the efficiency of the reducer and Table 4shows the test conditions. The sample was composedof a reducer and a hub as principal components. The
input shaft and the input shaft to the reducer werestructured separately, with the power between themconveyed by means of a splined joint. Matching thedeveloped IWM, the hub bearing was greaselubricated and the reducer was oil-lubricated. The oillubrication system in the reducer section was
configured so that a lubrication unit, which was
installed externally, circulated the lubricant to supply it
to the different parts of the reducer. The heater
installed in the circulation path kept the temperature of
the oil supplied at 70˚C.
The operating condition in which power wasconveyed from the input shaft of the reducer to the
output shaft, which simulates the power running of theIWM, was taken as the forward input condition. In
contrast to this, the operating condition in which power
was conveyed from the output shaft to the input shaftof the reducer, which simulates the regenerative

3.2 Motor section
Table 5 shows the test conditions for measuring the
motor efficiency. To match the developed IWM, themotor section was configured so that a lubrication unit,
which was installed externally, circulates the lubricant
to supply it to the rotor and stator of the motor.
Fig. 6 shows the results of the measurement of-24-

In-Wheel Motor System

motor efficiency. The horizontal axis shows therotational speed and the vertical axis is the motor
torque. The maximum efficiency of the motor section
was 94.7%.

Table 5 Test conditions for motor efficiency measurement
15,000 min-1

Maximum torque

45 Nm

LubricantAutomatic transmission oil

Lubrication system

Forced lubrication from the outside

Effeciency %
100

50

97

3.3 IWM
3.3.1 Test conditions

94

40

An efficiency measurement test was conducted in
the power running mode on the IWM system shown in
Fig. 2, which consists of a hub, a reducer and a motor.
In the power running mode, load was provided bymaking the built-in motor act as the power source andallowing an externally installed motor to absorb thepower generated. To measure the output torque, atorque meter was placed between the IWM and thepower-absorbing motor.
Table 6 shows the test conditions.

91

Torque Nm

Maximum rotational speed

88

30

85
82

20

79
76

10

73
70

0
0

2000

6000

10000

14000

Motor rotating speed min-1

Fig. 4 Reducer transmission efficiency (for forward input)
Table 6 Test conditions for IWM efficiency measurement
Effeciency %100

0Maximum rotational speed

97
-10

94

1365 min-1

Maximum torque

500 NmLubricantAutomatic transmission oil

Lubrication system

Built-in pump

Torque Nm

91
88

-20

-30

85

3.3.2 Test results

82

Fig. 7 shows the results of the measurement of theefficiency of the IWM under power running conditions.
The horizontal axis shows the wheel rotational speed
and the vertical axis shows the output torque. Table 7
shows the specifications of an EV comparable to atypical 1500cc-class compact automobile, and Fig. 8
shows the vehicle driving power calculated on thebasis of these specifications.
As shown in Fig. 7, the maximum efficiency of thedeveloped IWM was 90%, with a maximum power
output of 30 kW and a maximum torque of 500 Nm
being achieved. In addition, a power output of 26 kW
at a maximum rotational speed of 1365 min-1, whichcorresponds to a vehicle speed of 154 km/h, wasobtained. The results of calculations of driving powershown in Fig. 8 demonstrate that this is equivalent to
the power performance necessary to achieve arunning speed of 150 km/h on a road with a zerodegree incline for a common 1500cc-class compact
automobile propelled by two-wheel drive.

79
76

-40

73
70

-50
0

2000

6000

10000

14000

Motor rotating speed min-1

Fig. 5 Reducer transmission efficiency (for reverse input)
Effeciency %
100

50

97
94

40

Torque Nm

91
88

30

85
82

20

79
10

76
73
70

0
0

2000

6000

10000

14000

Motor rotating speed min-1

Fig. 6 Motor efficiency-25-

NTN TECHNICAL REVIEW No.79（2011）

Effeciency %
100

500

97

4.1 Test vehicle

94

Fig. 9 shows the appearance of the underbody of avehicle equipped with the developed IWM. The vehicle
is based on a commercially-available FF-vehicle, withthe rear wheels equipped with the IWM. The
suspension type is based on a double wishbone
system, with the auxiliaries being motorized andinstalled in the engine compartment with a battery
dedicated to powering them.
Fig. 10 shows the structure of the vehicle control
system.
The main constituent elements of the test vehicle
are the IWM system, a vehicle control unit (VCU) anda lithium-ion battery. The VCU gives the inverter
torque commands for the motors so that suitable
driving power is generated in the right and the left
IWMs independently in accordance with the driving
conditions on the basis of information from the driver,including acceleration, deceleration and corneringcommands. Following a torque command, the inverter
converts DC power from the battery into the AC power
required to drive the motor installed in the IWM.

450
400

Torque Nm

4. Test in Actual Car

550

91

350

88

300
85

250

82

200
150

79

100

76

50

73
70

0
0

200

400

600

800

1000 1200 1400

Wheel rotating speed min-1

Fig. 7 IWM efficiency

Table 7 Specifications of a typical 1500cc compact
vehicle class EV
Vehicle mass

1300 kg (including 110 kg for
the weight of 2 passengers)

Frontal projected area

2.6m2

Coefficient of drag

0.32

Rolling resistance coefficient

0.015

Driving power kW

80

Gradient of roadsurface
0 deg
2 deg

60

40

20

Fig. 9 IWM mounted in a vehicle (on the rear left wheel)
0
0

20

40

60

80

100

120

140

160

4.2 Temperature characteristics test

Vehicle speed km/h

The temperature characteristics of an IWM while
installed in a vehicle were assessed. The continuous
rating characteristic, one of the important performanceindexes of an IWM, is determined by the heat balance
between the heating of the IWM itself and the externalcooling action. In the assessment of the temperature
characteristics of this IWM that is air-cooled, testing inan actual car is effective for reproducing the runningwind. However, such a test poses difficulty in the
securing of the battery capacity necessary forcontinuous running and stabilizing the running load.
To solve this problem, we conducted tests using a
chassis dynamometer on which a blast duct was
installed as shown in Fig. 11. In accordance with theFig. 8 Vehicle driving power of a 1500-cc compact class EV-26-

In-Wheel Motor System

Brake

Brake

Ti
l
na
tio
ta
ro ed
re speAccelerator
pedal

Accelerationcommand

Brake
pedal

Turn
command

IWM
unit
DC
power
In-wheel motor
system information

Inverter

VCU

Indication andoperation panel

Deceleration command

Motor rotor angle
Motor rotor angle

Right and left wheel
torque commands
DC
power

AC
power

Battery

IWM
unit

Tire rotational
speed

In-wheel
motor system

Fig. 10 Composition of vehicle control system

speed of the rotation of the drum of the dynamometer,
which corresponds to the vehicle speed, air simulating
running wind was blown from the blast duct aroundthe IWM.
With the aim of supplying electricity in a stablemanner, an external power supply was used as thedrive power supply for the IWM instead of the onboard
battery.
In the test, the temperature of the motor coil section
and that of the internal lubricant were measured under
running conditions, with the vehicle speed and hillclimbing conditions used as parameters.
In this test, the allowable upper limit of the coiltemperature was set at 180˚C from the viewpoint of
the heat resistance of the coil insulation film. In
addition, with the aim of preventing a reduction in the
durability of the bearings and reducer parts used in
the IWM, the allowable upper limit of the lubricant
temperature was set at 120˚C.
As shown in Table 8, the test was conducted under
the conditions of running resistance and a runningwind speed that correspond to a case where thevehicle is running at a speed of 70 km/h on a slopewith a gradient of 2 degrees. Fig. 12 shows the test
results. Both the coil and the lubricant reachedsaturation temperatures that were below their
allowable upper temperatures. The lubricant exhibiteda smaller difference between the allowable upper limit
temperature and the saturation temperature reached.
Considering that this temperature difference is about
10˚C, the assessment shows the possibility ofcontinuous running under the load conditions shown in
Table 8 if the outside air temperature is below 35˚C orso.

Table 8 Conditions for the temperature characteristics test
Equivalent to 70 km/h
（655 min-1）

Vehicle speed(Wheel rotational speed)Wind speed

Equivalent to 70 km/h

Hill climbing resistance

Equivalent to an incline of 2 degrees

Running air temperature

25˚C

130

Blast duct
Temperature ˚CCoil temperature

Running
wind

110

90

Lubricant temperature
70

50
0

20

40

60

Time minFig. 11 Vehicle test with chassis dynamometer

Fig. 12 Results of temperature characteristics test

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80

NTN TECHNICAL REVIEW No.79（2011）

Reference

As shown in Fig. 8, the drive power for a carrunning at a speed of 70 km/h on a slope with a 2degree gradient is nearly equal to the driving power
for a car running at a speed of 100 km/h on a levelroad surface. If, for simplification of assessment,variation in the efficiency of the IWM due to variationin operating conditions is ignored, we may conclude
that NTN's newly-developed IWM system is capable ofcruising at 100 km/h on a level road surface (zerodegrees gradient) because increased car velocity
results in a greater forced convection cooling effect.

1) Imayanagita, A. et al., "Development of In-wheel Motor
System for Large-size Bus using 22.5-inch Wheel
Mounted Motor," Toyo Denki Review, No. 113, pp. 914, 2006.
2) Tahara, Y. et al., Development of the In-Wheel Motor
System, Transactions of the Society of Automotive
Engineers of Japan, No. 131-06, 2006.
3) Okamoto, Y. et al., Development of High Efficiency Inwheel Motor System for using 20-inch Wheel Mounted
Motor, Toyo Denki Review, No. 117, pp. 12-17, 2008.
4) Kaneko, Y. et al., Development of Compact and HighPower In-Wheel-Motor unit, Transactions of theSociety of Automotive Engineers of Japan, No. 7-09,
2009.
5) Murata, S., Development of In-Wheel Motor Drive Unit,
Transactions of the Society of Automotive Engineers ofJapan, No. 28-10, 2010.
6) Akaho, D., Development of Vehicle Dynamics Control
System for In-Wheel-Motor Vehicle, Transactions of
the Society of Automotive Engineers of Japan, No.
120-10, 2010.
7) Suzuki, M. et al., NTN Technical Review, No. 73, pp.
56-59, 2005.
8) Suzuki, M. et al., NTN Technical Review, No. 75, pp.
46-52, 2007.
9) Morozumi, M., Theory and Design Calculation Methodof Planetary Gears and Differential Gears, Nikkan
Kogyo Shimbun Ltd., pp. 1-6, 1989.

5. Conclusion
Assuming installation in the two rear wheels of a
1500cc-displacement compact automobile, wedeveloped the NTN-IWM System in which a control
system driven by sensor-generated information isincorporated into each IWM, which consists of acycloid reducer and a motor.
The efficiency of each element of the IWM system
was assessed. The highest efficiencies achieved were
98% in the forward direction and 97.5% in the reverse
direction for the reducer section, 94.7% for the motor
section, and 90% for the IWM. In addition, we
obtained power equivalent to that which allows acommon 1500cc-class compact automobile with twowheel drive to run at a speed of 150 km/h on a roadwith no incline.
Furthermore, we conducted a temperature
characteristic test using a chassis dynamometer with
the newly-developed IWM system onboard a vehicle,
confirming that it is possible for the vehicle to run
continuously at a speed of 100 km/h on a road withzero-degree incline in ambient air temperatures below
35˚C.
We are continuing road tests using a test vehicle
with the objective of further improving reliability andcommercializing our IWM system as early as possible.

Photo of authors

Yuich ITOH

Kayo SAKAI

Yusuke MAKINOEV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Control System
Engineering Dept.

-28-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Article ]Development of In-Wheel Motor System
for Electric Commuters

Tetsuya YAMAMOTO*
Aiko ISHIKAWA*
Wataru YAMADA**NTN has developed an In-wheel Motor System for commuter vehicles that
could become a stepping stone that promotes the spread of electric vehicles(EV). The in-wheel motor unit that was developed can be installed in 14-inch
wheels. In the test EV, it was shown that space in the car was able to be
utilized effectively. This paper introduces the In-wheel Motor system thatwas developed and the test EV.

can be recharged with late night electricity
3) Commuters intended for use in short-to-medium
range distances from residences (daily travel ofless than 40 km)

1. Introduction
In general, two problems exist in promoting thewidespread use of electric vehicles (EVs). One is the
reduced cruising range when compared to that ofordinary automobiles with internal combustion engines
and the other is the higher initial costs of batteries.
These two problems are expected to be solved in the
future as technologies advance. For example, to solve
the first problem of a reduced cruising range, what is
important is not only technological progress for
improved battery performance but also ubiquitous
availability of battery-recharging infrastructure. For
these problems, it is hoped that efforts are made byboth the government and the private sector in order torealize the widespread use of EVs 1).
Against this backdrop, the Ministry of Economy,
Trade and Industry planned the Next-Generation
Vehicle Strategy 2010 2) with the aim of spreading theuse of next-generation vehicles. On the basis of theroadmap shown in the METI strategy, focused support
for the development of infrastructure, such as thecreation of "electric vehicle/plug-in hybrid vehicle(EV/PHV) towns," will be undertaken.
Vehicles that will be desired in such EV/PHV towns
in the future are thought to be of the following types.

The services required of the commuter vehicle type
in item 3 could be met to a certain extent by smallsized one-person commercial delivery vehicles if
equipped with a large cargo compartment. To achieve
the widespread use of EVs, however, creating amarket intended for general users, not just commercial
users, is essential. A fiscal 2009 survey, focusing on
the uses of vehicles with small displacement engine,
showed that about half of the users mainly used theirvehicles for shopping and picking up family members.
Needs for electric commuters are expected to be highin similar applications, as well.
Taking this into consideration, the electric commuter
could be a vehicle that meets a wide variety of needs
and fulfills the promise of extensive popularization if
measures are taken to seat two or more passengers
while maintaining the compact size of this type of
vehicle.
Furthermore, designing a vehicle for use over short
to medium distances makes it possible to reduce theweight load of the battery and thus the cost. A case ofdevelopment of EVs for which costs were reduced
from this point of view has been reported. 4)
Securing space to accommodate luggage andsimilar cargo is also important for commuter vehicles.

1) Fixed route buses that can be operated andrecharged on schedules
2) Delivery vehicles that are used in limited areas and**EV System Division Driveline System Engineering Dept.
**EV System Division Control System Engineering Dept.

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NTN TECHNICAL REVIEW No.79（2011）

Table 1 Target specifications of the in-wheel motor system

To power EVs, motors replace the internal combustion
engines of conventional automobiles. Compared toordinary EVs, in which power is conveyed to thewheels via the drive train, in-wheel EVs with motors
installed in their wheels will allow for increases inpassenger and luggage space without enlarging thevehicle size.
For this reason, taking note of two-seatercommuters for short-to-medium distance travel, NTN
developed an in-wheel motor system for use in electriccommuters. This system is composed of in-wheel
motors that can be mounted in the wheels and an
inverter for motor control, and was designed with theaim of increasing freedom in vehicle body design andusable cabin space. This paper reports on the conceptbehind this newly developed product, its
characteristics, and a vehicle equipped with thissystem.

In-wheel motors

Cooling system

Hub

Maximum output

2kW

Maximum torque

132Nm

Suitable wheel size

14 inchIn-wheel motor

Air-cooled

Inverter

Water-cooled

Battery voltage

96V

Maximum vehicle speed

60km/h

Housing

Suspension
Motor andmounting part reducer sections

2. Specifications of this in-Wheel
Motor System
Under current legal restrictions in Japan, the twoseater four-wheeled vehicle covered by this paper falls
under the category of a small-displacement engine
vehicle, assuming that the restrictions on its vehicle
size and prime mover output are met. In Europe, aspecial standard applied to the two-seater commuteris already in effect.
Taking the current situation and trends of legal
restrictions into account, the specifications for an inwheel motor system were defined as shown in Table
1.
This in-wheel motor was designed so that it could
be placed inside a 14-inch wheel. Fig. 1 shows theappearance of the in-wheel motor that was developed.
This in-wheel motor is composed of a motor section, areducer section and a hub section. To reduce the sizeof the entire assembly, the mounting part to thesuspension was incorporated into the housing.
Assuming that the commuter does not run on
expressways but rather on urban streets in mostcases, its highest speed was set at 60 km/h, the legal
upper limit on ordinary roads. Fig. 2 shows the resultsof measurement of the efficiency of the developed inwheel motor. Based on the upper limit of 60 km/h forurban the target was to have the highest efficiency at
the frequently used low-to-medium speed range (2040km/h) and this was achieved per Fig. 2.

Fig. 1 Exterior view of the in-wheel motor

Effeciency %
140

87
85

120

Wheel Torque Nm

83
81

100

79

80

78
60

76
74

40

72
20
70
69

0
0

10

20

30

40

50

Vehicle speed km/h

Fig. 2 In-wheel motor efficiency

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60

Development of In-Wheel Motor System for Electric Commuters

in-wheel motor. This space can be used as a luggage
compartment when luggage-carrying capacity is
considered important. In this vehicle, a Li-ion battery is
installed below the seats in the center of the body.
When emphasis is placed on the cruising range, theusable space described above can be used to carryadditional batteries. Optimized vehicle frame andsuspension system will help realize more efficientutilization of the space within the vehicle.

3. Test Vehicle
A test vehicle equipped with this in-wheel motor
system was prototyped. Table 2 shows the vehicle
specifications.
Fig. 3 shows a diagram of the configuration of thetest vehicle and Fig. 4 shows the appearance of theunit when installed on the vehicle.
The vehicle control unit (VCU) determines thetorque required for the motor on the basis of stroke
information from the accelerator operated by the driverand sends torque commands to the inverter. In turn,the inverter converts the DC power from the battery
into the AC power necessary to drive the right and left
motors, enabling them to generate suitable driving
power according to the driving conditions.
A running test was conducted on this test vehicle instraight-ahead driving, cornering, and slalom
conditions, and the achievement of smooth andpowerful acceleration, which is characteristic of EVs,
and of satisfactory stability drive control wereconfirmed.
Fig. 5 shows a photograph of the space behind theseats of the test vehicle. A usable space with avolume of about 40 L was created between the righthand and left-hand suspensions each supporting theTable 2 Specifications of the test EV
Vehicle weight

315 kg (not including the body cowl)

Vehicle size (L×W×H)

2580mm×1390mm×1350mm

Tire size

155/55R14

Drive wheels

Two rear wheels

Maximum output

2kW×2

Maximum output torque132Nm×2

BatteryType

Li-ionCapacity

3.1kWhFig. 4 Test EV (above) and mounted in-wheel motor(below)

DC power
AccelerationcommandBatteryMotor rotor Motor
angle(Acceleration stroke)

In-wheel motor
systeminformation
AC power

InverterIndi Control
cators panel

VCU
Torque command toright and left wheels

AC power

DecelerationcommandMotor rotor
angle

Motor

In-wheel motor system

Fig. 3 Schemetic of EV with in-wheel motor system

Fig. 5 Luggage space behind seats-31-

NTN TECHNICAL REVIEW No.79（2011）

4. Conclusion
Taking note of two-seater commuters that will
provide a springboard for the popularization of EVs,
NTN has developed an in-wheel motor system thathas an advantage in creating cabin space. The
running test on the test vehicle equipped with the
NTN's newly developed in-wheel motor system has
confirmed the smooth and powerful acceleration
performance associated with EVs as well as
satisfactory stability in drive control.
We will continue our development efforts toward theimprovement of performance and reliability with thegoal of contributing to the popularization of two-seater
EV commuters.

Reference
1) Ministry of Economy, Trade and Industry, Guidebook
for Installing Charging Facilities for Electric Vehicles
and Plug-In Hybrids, 2010.
2) Ministry of Economy, Trade and Industry, NextGeneration Vehicle Strategy 2010, 2010.
3) Japan Automobile Manufacturers Association, Reporton Investigation of the State of Light Car Utilization,
2010.
4) Shimamura, K., Hirose, H., Development of Next
Generation EV Concept "C･ta" (third report),
Transactions of the Society of Automotive Engineers ofJapan, No. 51-1020105142, 2010.
5) European Commission, 558340 DT Two- or Threewheel vehicles and Quadricycles Impact Assessment,
SEC (2010) 1152, 2010.

Photo of authors

Tetsuya YAMAMOTO

Aiko ISHIKAWA

Wataru YAMADAEV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Control System
Engineering Dept.

-32-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Paper ]One Motor Type Electric Vehicle Drive System

Fumihiro ISOBE*
Yoshinori ITAKURA*
Yusuke OSUMI*
Guodong LI**
In order to increase the cruising range per charge of one-motor type electric vehicles
by improving power train efficiency, a one-motor type electric vehicle drive system thatutilizes a 2-speed transmission has been successfully developed. To reduce the size
and weight of the transmission while keeping high torque capability, a two-way roller
clutch has been adopted as the shifting mechanism. This paper describes thecomposition of the developed system, the structure of the transmission, the principle of
the gearshift, and the results of gearshift tests on a test bench and a test vehicle.

controller for controlling the shifting action. This paper
reports on the structure of the transmission, theprinciple of gear shifting, the results of shifting tests on
a test bench and the results of shifting tests on an
experimental vehicle.

1. Introduction
In response to global warming and problems
associated with energy, the development of hybrid
and electric vehicles has been accelerating. In thesecircumstances, mass production of electric vehicles ofthe one-motor type, which are driven by power
conveyed from a single onboard electric motor to theright and left wheels via a differential and a driveshaft,
has already begun.
Many of the commercialized one-motor-type electricvehicles are equipped with reducers that lack speedshifting functionality, and their maximum torque andmaximum rotational speed are determined by atradeoff between the acceleration performance andthe maximum speed required with the reduction ratioof the reducer. Setting the reduction ratio to a lower
value while increasing the maximum motor torqueresults in the size of the motor itself becoming larger.
In contrast, setting the reduction ratio at a higher value
while increasing the maximum rotational speed of thedrive motor results in an increase in the power loss of
the reducer, and hence in a reduction in efficiency.
In response to these problems, with the aim of
improving the electricity consumption ratio of onemotor-type electric vehicles, for EV-use, NTN has
developed a two-speed transmission and a drive
module composed of a smaller, lighter drive motor anda differential, along with an NTN one-motor-EV drive
system integrating a drive-motor inverter and a shift

2. Development Concept
2.1 Effectiveness of two-speed transmissions
As mentioned above, the performance required of adrive motor with a fixed reduction ratio is determined
specifically by the reduction ratio, the output torquethe vehicle requires and the vehicle speed. When adrive motor with a larger maximum torque is used withthe reduction ratio set to a lower value this leads to anincrease in the size and mass of the motor. For thisreason, setting the reduction ratio at a higher value iseffective for reducing motor weight. 1) In this case,however, it is necessary to increase the maximum
rotational speed of the drive motor. This results ingreater power loss in the reducer, leading todecreased transmission efficiency.
Fig. 1 shows the concept behind the advantage ofintroducing a two-speed transmission system.
Switching appropriately between the high reduction
ratio gear and the low reduction ratio gear, for either
low or high speeds respectively, allows the maximum
torque of the drive motor to be set at a lower value
than that of a fixed-reduction-ratio motor without
increasing the maximum motor rotational speed. This**EV System Division Driveline System Engineering Dept.
**EV System Division Control System Engineering Dept.

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NTN TECHNICAL REVIEW No.79（2011）

makes it possible to reduce the motor weight andincrease the transmission efficiency at the same time.
The efficiency of the drive motor and that of theinverter vary depending on the rotational speed andtorque, and the ranges of rotational speed and torquein which operation at higher efficiencies is allowed arelimited. As shown in Fig. 1, adopting a two-speedtransmission system enables the range in which
operation at higher efficiencies is possible to beextended. It is possible, from this viewpoint also, toincrease the system efficiency further and to reduce
the electricity consumption necessary to drive avehicle.
In the past, the effectiveness of the two-speedtransmission scheme has been verified several times
by means of simulation and other methods, and it has
been reported that this scheme is capable of reducing
the rate of electric power consumption by over 10%
compared with a fixed reduction ratio scheme. 1)-3)

speed stage is selected in the low-to-medium speed
range when great driving power is needed. First, thereduction ratio of the second speed stage is selected
temporarily, depending on the allowable maximum
rotational speed of the motor selected and theplanned maximum vehicle speed. On the basis of this
temporarily-selected second speed stage reduction
ratio, the first speed stage is selected temporarily sothat adequate drive power in the low-to-medium speed
range can be achieved while reducing the size of thedrive motor and that smooth shifting between the first
and the second speed stages is achieved.
Adjustments of the reduction ratios, temporarily
selected as described above, were made to increasesystem efficiency on the basis of the efficiency
characteristics of the inverter and the transmissionsection, with the reduction ratios being set as shownin Table 1.
Table 1 Development specifications

High reduction ratio(Torque at low-speeds emphasized)

SystemOutput torqueLow reduction ratio(High-speed running emphasized)

Output torqueHigh-efficiency range

Vehicle speed

Motor
Drive
module Transmission

60 kW

Maximum output torque1000 Nm

Maximum applicable vehicle speed 150km/h
Type
IPM
First
speed
Second(including the final reduction ratio)
speed

Reduction ratio

1/9.9
1/5.6

Vehicle speed

Two-stage reduction ratio
Increased drive powerin low-speed range

Output torqueMaximum output

3. Structure and Operating Principleof Transmission
3.1 Features of typical transmission mechanisms

Expanded high-efficiency range(Increased efficiency)

Since the change from a fixed reduction ratio
scheme to a two-speed shifting scheme leads to anincrease in the transmission mass, it is important for
the gear change mechanism to be lightweight andcompact. This is to ensure that its weight does notoffset the decrease in weight achieved by reducing thedrive motor weight. In addition, it is also important forsuch a shifting mechanism to be a highly efficienttransmission from the viewpoint of improving the rateof electricity consumption.
Typical automotive automatic transmissions include
automatic transmissions (AT), continuously variabletransmissions (CVT), dual clutch transmissions (DCT)
and automated manual transmissions (AMT). The
features of each of these transmissions are described
below.

Expanded high-speed
running range
Vehicle speed

Fig. 1 Expansion of the high-efficiency range by
employing a two-speed transmission scheme

2.2 Development specifications
A small-sized automobile with a 1000cc-classdisplacement was set as the target of development,
and the specifications of the vehicle to be developed
were defined as shown in Table 1.
The reduction ratios were defined using thefollowing process. In the present development work,
the major purpose was to reduce the power loss in thetransmission associated with an increase in therotational speed. For this reason, the second speed
stage is selected when running at relatively low speed
and load or running at high speed, while the first(1) AT
Recent progress in electronic hydraulic control has
made it possible to change gears with little shock.
However, the necessity of placing a torque converter-34-

One Motor Type Electric Vehicle Drive System

between the prime mover and the transmission is a
disadvantage in terms of size and efficiency. The
hydraulic pump and hydraulic circuit necessary tocontrol the oil pressure are also factors that increase
the size and cost.

3.2 Structure of the transmission to be developed
In the present development work, the characteristics
required of a two-speed transmission were set as
follows.
(1) Lightweight and compact(2) High transmission efficiency(3) Quick shifting(4) Non-use of oil pressure in shifting control(2) CVTThe most common type of automotive CVT is themetal-belt type. This CVT achieves smooth shifting
without shock. However, the necessity for two pulleys,
one each on the primary and the secondary sides, andthe mainstay structure consisting of a torque converter
as in an AT generally tend to make CVTs larger insize and mass and more expensive than othertransmissions. In addition, the transmission of powerby means of friction between the pulley and the metal
belt requires large belt clamping force, and thegeneration of high oil pressure for this purpose
increases power loss, which in turn lowers thetransmission efficiency. An effective use of a CVT is touse it in combination with a prime mover that canoperate in a wide range of gear change ratios, which
is a notable CVT feature.

The conventional shifting mechanisms described in
Section 3.1 were judged to not be capable of meeting
all these requirements, so a scheme using a roller
clutch in the shifting mechanism was devised.
Fig. 2 shows the structure of the newly developed
drive module. The transmission consists of three
shafts, namely, an input shaft that conveys powerfrom the motor, a differential connected to the output
shaft of the reducer (a constant velocity joint), and acountershaft that connects the input shaft and thedifferential. The first-speed and the second-speed
gear trains are of the constant mesh type, and thetorque output from the drive motor is transmitted to thecountershaft via one of the two gear trains and fromthere to the differential and the constant velocity joint.
The countershaft is equipped with the first-speed
and second-speed gears, each of which incorporates
a roller clutch, and the roller clutch on/off control
mechanism is placed between them. In thistransmission, a separately-installed motorized speedchange actuator operates the roller clutch on/off
control mechanism, thereby engaging anddisengaging the two roller clutches to switch between
the first and second speeds.

(3) DCT
Recently, DCTs have been adopted for use ininternal combustion engine vehicles in an increasing
number of cases. Although a DCT allows quick gearchanges, it needs a dual clutch unit consisting of two
wet multi-plate disk clutches or dry single-plate disk
clutches and a clutch changeover actuator between
the prime mover and the transmission. Moreover, themain body of the transmission requires two or more
input shafts, which increases the part count.

Roller clutch on/off
control mechanism(4) AMT

Drive motor

Input shaft

Based on a manual transmission (MT), an AMT has
a basic structure consisting of a main clutch (usually adry single-disk clutch) between the prime mover andthe transmission and two or more actuators with whichto perform shifting operations. An AMT allows anautomatic transmission to be formed with a relatively
simple structure, and it excels in terms of efficiency
since it is based on an MT. However, the installationof a main clutch results in larger transmission
dimensions, and the need to cut off power by meansof the main clutch in order to change gears can resultin a longer shift time when compared to othertransmissions.

Second-speed
gear train

First-speed
gear train

Countershaft

Constant velocity joint
Roller clutch
Differential

Fig. 2 Basic structure of newly developed motor module-35-

NTN TECHNICAL REVIEW No.79（2011）

Some advantages of the roller clutch over otherclutch mechanisms, such as wet multi-plate disk
clutches, are: (1) the large torque transmission
capacity allows for the size (outside diameter) to be
reduced; (2) the structure is simple; and (3) the dragtorque in the disengaged condition is small. The firstand second advantages make it possible to place aclutch section that has sufficient load capacity within
the bore of the gear to form a small, lightweighttransmission. Since it is based on a parallel biaxial
gear reduction scheme like that of an MT, thistransmission has higher efficiency over a wide range
of operating conditions. The third advantage describedabove allows for high efficiency to be achieved if it isused in a two-speed transmission.
The roller clutch can be engaged even if the outerring and the inner ring of the clutch to be engaged are
not perfectly synchronized with each other in terms of
their respective rotational speeds. For this reason, thesynchronous action of the drive motor alone canchange speeds quickly and with less shock comparedto the friction-based synchronizers used in DCTs and
AMTs. Furthermore, the elimination of a synchronizer
also eliminates the necessity of installing a main
clutch to cut off power between the motor and thetransmission, contributing to a reduction in the size
and weight of the transmission.
Using a roller clutch in the shifting mechanism
makes it possible to achieve all four of the necessaryfunctions described above and realize a two-speedtransmission suitable for electric vehicles.
A trial calculation carried out with the conditions thatthe maximum output, maximum output torque,maximum vehicle speed, and maximum motor
rotational speed were equivalent shows that the massof NTN's newly developed drive module combining thetransmission drive motor is smaller by about 20% than

Return spring

Rollerthat of a drive module with a fixed reduction ratio.

3.3 Gear shifting principle
Fig. 3 shows the operating principle of the roller
clutch. The torque transmitting section of a roller
clutch consists of an inner ring with two or more cam
faces on the outer circumference, an outer ring with acircular rolling face on the inner circumference, rollers
installed between the inner ring cam face and theinner circumferential face of the outer ring, and thecage that holds the rollers at the same intervals in thecircumferential direction. This design forms a wedgeshaped space between the inner ring cam face andthe inner circumferential face of the outer ring. A
return spring that is installed between the inner ringand the holder allows the energizing force, acting in
the direction that negates the gripping of the roller inthe wedge-shaped space, to act on the cage.
Fig. 4 shows in detail the structure around thecountershaft. The roller clutch on/off control
mechanism consists of two friction disks, one each forthe first and second speeds, and a friction disk
retainer. Each friction disk is fitted into the counterpart
cage in such a manner so as not to allow relativerotation but to allow axial movement. The operation of
the transmission actuator causes the shift fork to be
driven axially, and the shift fork in turn causes thefriction disk retainer and the friction disk to moveaxially. By moving the friction disk toward the roller
clutch with which the friction disk is to be engaged andallowing the friction disk to come in contact with theouter ring, a frictional force that transmits the rotationof the outer ring to the friction disk is generated, andthe rotation of the outer ring is transmitted to the cage
via the friction disk. The transmission actuator iscontrolled so that frictional force sufficient to allow theouter ring to rotate together with the cage is produced.

Outer ring

ClearanceCage

Cam face

Inner ring(a) Neutral(b) Engaged

Fig. 3 Roller clutch operating principle-36-

One Motor Type Electric Vehicle Drive System

When frictional force does not act between the outerring and the friction disk, the cage and the roller are
held at the central position (the neutral position) of the
cam face by the force of the return spring. In thiscondition, clearance occurs between the roller and theouter ring or inner ring, and the inner and the outerring are put into a neutral state allowing relativerotation. On the other hand, when a frictional force
larger than the force of the return spring acts on theouter ring and the friction disk, the relative rotation of
the inner and the outer ring causes the cage, rotating
together with the outer ring, to move the roller into thewedge-shaped space (the position of engagement) as
shown in Fig. 3 (b), with the inner and the outer ring
engaged with each other via the roller.
This roller clutch is a so-called two-way clutch that
can transmit torque both clockwise andcounterclockwise, so it is capable of transmitting
torque not only in the power running mode but also in
the regenerative mode.

To the transmission actuator

Friction disk retainer
RollerSecond speed
inner ring

4. Shifting Performance
4.1 Shifting control
In a roller clutch, the engagement of the roller ismaintained as long as the transmission torque isabove zero even if the holder does not push the roller
into the wedge-shaped space. For this reason, theengagement of the roller clutch is maintained as longas torque transmission is continued, even after thecontact of the outer ring with the friction disk by meansof the action of the roller clutch on/off control
mechanism is released. For this reason, shifting
operations are performed in the following order:(1) First, the shift fork is moved from the current speed
stage position to neutral to stop contact between
the outer ring and the friction disk.
(2) The motor torque is lowered to nearly zero to stopthe engagement of the roller clutch.
(3) The drive motor is accelerated or decelerated until
the rotational speeds of the inner ring and the outerring of the roller clutch of the next shifting stage
match (synchronization).
(4) Concurrently with the start of synchronization, theshift fork is made to start moving toward the next
shifting stage. In this step, the speed of thesynchronization is adjusted, so that the completionof the synchronization does not lag behind thecompletion of the movement of the shift fork, thatis, the time of generation of the frictional force
between the outer ring of the roller clutch of thenext shifting stage and the frictional disk.
(5) After completing synchronization, the movement of
the shift fork to the next shifting stage iscompleted, with the outer ring of the roller clutch of
the next stage allowed to come in contact with thefriction disk.
(6) Entering the drive motor, torque again causes theroller clutch of the next shifting stage to beengaged, with the torque transmission started andthe shifting operation completed.
The information on the input/output rotational speed
necessary for the control of shifting is obtained from arotational angle detection sensor built into the drive
motor on the input side and from the min-1 detection
sensor installed in the differential section on the output
side. Information about the position of the shift fork is
detected by a position sensor attached to the shift
fork.

Shift fork

Return springFirst speed
inner ring

RollerSecond speed
outer ring
Second speed
holderSecond speed
friction disk

First speed
outer ringFirst speed friction disk
First speed
holderFig. 4 Detail of countershaft section

4.2 Bench shifting test
4.2.1 Testing equipment
To carry out the bench test, a dummy differential
integrating the ring gear and the shaft (the output
shaft), instead of the original differential, was-37-

NTN TECHNICAL REVIEW No.79（2011）

incorporated into the drive module specimen. Load
was provided by making the motor, built into the drive
module, act as the power source and allowing an
externally installed motor to absorb the power
generated. To measure the output torque, a torquemeter was placed between the specimen and thepower-absorbing motor.

was about 0.23 sec and for downshifting was about
0.27 sec.
The reason why downshifting takes more time isthat the torque of the drive motor is lowered to zero todisengage the roller clutch and the standby time tomaintain this condition is set. As described in Section
4.1, the roller clutch is not disengaged as long as thetransmission torque does not become zero. For thisreason, when the torque transmission direction
immediately before shifting and the direction in which
the drive motor torque is generated in thesynchronization are the same as in the case ofdownshifting, it is necessary to lower the torque of thedrive motor to zero once and to wait for thedisengagement of the roller clutch.
In upshifting, on the other hand, the torquetransmission direction immediately before shifting andthe direction in which the drive motor torque isgenerated in the synchronization are opposite to eachother even in the same accelerating condition, with thesynchronization itself acting to end the engagement of
the roller clutch. This eliminates the time waiting for
disengagement. For this reason, downshifting takes
0.04 sec longer than upshifting does. Conversely,
when the shifting occurs in a state in which the torqueto decelerate the vehicle is output from the drive
module as in the case when the regenerative brake ison downshifting takes a shorter time than upshiftingdoes.
The shifting test was conducted on a bench up to a
maximum vehicle speed of 80 km/h with a tire
diameter of 600 mm and in a drive motor torque rangeof 20 Nm to 60 Nm and confirmed that smooth shiftingis achieved without the occurrence of excessive
torque.

4.2.2 Test method
To shift upward, the first speed stage is selected
beforehand. To shift downward, the second speed
stage is selected beforehand. The initial condition isdefined as follows: the external motor for the power
absorption purpose is operated at a fixed rotational
speed through the rotational speed control and thedrive motor produces constant positive torque. The
positive torque in this case is that which accelerates
the vehicle in the forward direction.
Beginning with this condition, the drive motor andthe transmission actuator are controlled in thesequence described in Section 4.1 to change speeds.

4.2.3 Results of the test
Figs. 5 and 6 show the results of upshifting anddownshifting from an output rotational speed of 441
min-1 (equivalent to a vehicle speed of about 50 km/h
with a tire of 600mm diameter) at a drive motor torqueof 60 Nm.
The shifting time is defined as the time required
during which the output torque from the drive module
is lowered to nearly zero for shifting and, after thecompletion of shifting, the drive module starts
outputting torque again. Under the conditions of adrive motor torque of 60 Nm and an output rotational
speed of 441 min-1, the time required for upshifting

Moduleoutput torqueTime waiting forclutch disengagement

Shifting time: 0.23 sec
Synchronization time

Shifting time: 0.27 sec

500

4000

400

3000

300

2000

200

1000

100

0

0.4

Drive motor
rotational speed

0.6

0.8

1

Time (sec)

1.2

1.4

6000

600

5000

500

4000

400

3000

Module output
torque (Nm)

2000

200

1000

100

0
1.6
Drive motor
rotational speed

Module output
rotational speed

Fig. 5 Results of upshifting test

1.8

2Time (sec)0
4.4
Module output
rotational speed

2.2

Fig. 6 Result of downshifting test

-38-

300

Module output torque (Nm)5000

Drive motor rotation speed (min-1)
Module output rotation speed (min-1)

600

Module output torque (Nm)

Drive motor rotation speed (min-1)
Module output rotation speed (min-1)

Synchronization time
6000

One Motor Type Electric Vehicle Drive Systemtime. Under these circumstances, the moving time ofthe shift fork is the major factor for determining theshifting time, and improving the performance of theshift actuator responsible for the shift fork movement
could bring about further reduction in the shifting time.
An irregularity of about 0.02 sec that is observed in
the shifting time during upshifting or downshifting isdue to the irregularity in the time lag from thereapplication of drive motor torque to the engagementof the next shift stage roller clutch. That is, due to theirregularity in the time required to move the shift fork tocause the roller clutch outer ring to contact thefrictional disk and then, due to the torque input fromthe drive motor, for the roller of the roller clutch tomove from the center position of the cam face (neutral
position) to the wedge space (engaged position).

The effect of the vehicle speed and the drive motor
torque on the shifting time is shown in Figs. 7 and 8,
with the former showing upshifting and the latter
downshifting. In the vehicle speed range of 20 km/h to
80 km/h, upshifting took about 0.22 sec to 0.24 sec
while downshifting took about 0.26 sec to 0.28 sec.
This shows that the shifting time is nearly constantregardless of the vehicle speed. The shifting time isdefined as the time required for the torque output fromthe drive module to become zero and to rise again.
This time is determined by the time required for theroller clutch at the current shift stage to bedisengaged, the time for the shift fork to move, andthe time lag between the reentry of the drive motor
torque and the engagement of the next shift stage
roller clutch. With these times not dependent on thevehicle speed and output torque at the time of shifting,the shifting time is nearly constant. The time for therotational speed synchronization depends on thevehicle speed. However, with the synchronization able
to be completed within the shift fork moving time forthe entire speed range, the time for the rotational
speed synchronization does not affect the shifting

5. Onboard Test
The actual vehicle test drive was conducted with theaim of assessing the shifting performance under
actual usage conditions that are difficult to reproduce
in a bench test.

5.1 Test vehicle configuration
0.30

Drive motor
torqueShifting time (sec)

0.28

Fig. 9 shows the drive module mounted on thevehicle and Fig. 10 the major system configuration for
the test vehicle.
A commercially available FF vehicle was utilized forthe test. The engine and transmission were removed,with the engine compartment modified so as toaccommodate a drive module and its inverter.
The drive motor was water-cooled, with the coolingwater being circulated between the radiator, inverter,and drive motor by a newly-installed electric pump.

20 Nm
40 Nm
60 Nm

0.26
0.24
0.22
0.20
0.18
0.16
20

30

40

50

60

70

80

Vehicle speed (km/h)

Fig. 7 Time needed for upshifting

Vehicle frontTwo-speedtransmission

Drive motor
0.30

Shifting time (sec)

0.28
0.26
0.24
0.22
0.20
20 Nm
40 Nm
60 Nm

Drive motor
torque

0.18
0.16
20

30

40

50

60

70

80

Vehicle speed (km/h)

Differential

Fig. 8 Time needed for downshifting

Constant velocity joint

Fig. 9 Module mounted in the test vehicle-39-

NTN TECHNICAL REVIEW No.79（2011）

ForwardRadiator

Differential
Output rotational
speed information
Accelerator
AC power
Accelerator stroke
information

VCU

Drive
motor

Shifting
controller

Battery
Shift actuator
control signal

Inverter
DC power

Drive motor control signal

Drive motor rotationa
speed information
Electric pumpTwo-speed transmission
Cooling water circulation path

Shift fork position information

Fig. 10 Drive system for a one-motor-type electric vehicle

The drive battery was placed in the space that had
been occupied by the gasoline tank.
A shifting controller was installed to control theshifting of the drive module, based on accelerator
travel, drive motor rotational speed, output rotational
speed, and shift fork position information.

Clutch disengagement
wait time

Shifting time: 0.3 sec
Synchronization time

5.2 Shifting tests
5.2.1 Test method
With the shift controller installed in the cabin, thetests were conducted by manipulating the controller
manually so that speed could be changed at any time.
The tests were conducted at vehicle speeds between
20 km/h and 60 km/h.

300

4000
Drive motor
rotation speed

200

3000

2000
100
1000

0

0.6

0.8

Wheel output
torque5.2.2 Results of the tests

Wheel output torque (Nm)

Drive motor rotational speed (min-1)
Module output rotational speed (min-1)

5000

1

1.2

0

Module output
rotational speed

Time (sec)Fig. 11 Results of actual vehicle upshifting test

Figs. 11 and 12 show the results of upshifting anddownshifting at a vehicle speed of about 50 km/h.
In these tests, the output torque from the drive
module was measured by a wheel torque meter
mounted on the left wheel. The wheel output torquedata shown in Figs. 11 and 12 are the measured
values. Since they were measured on a single wheel
only, they show about half of the actual torque output
from the module.

Clutch disengagement
wait time

Shifting time: 0.29 sec
Synchronization time

4000

300

Drive motor
rotation speed
200

3000

2000
100
1000

0
0.6
Wheel output
torque0.8

1Time (sec)01.2Module output
rotational speed

Fig. 12 Results of actual vehicle downshifting test

-40-

Wheel output torque (Nm)

Drive motor rotational speed (min-1)
Module output rotational speed (min-1)

5000

One Motor Type Electric Vehicle Drive System

In the future, a reduction of the clutch
disengagement waiting time in upshifting, set in theactual vehicle test, will be implemented by improving
shift control through repeated in vehicle testing, until
the shift time is shortened to that of the bench test.

Table 2 shows the shifting times required forupshifting and downshifting at 25 km/h, 50 km/h and
65 km/h.
Table 2 Time needed for gear shifting in the test vehicle
Vehicle speed (km/h)

25

50

65

Upshifting

0.29

0.30

0.30

0.29

0.29

0.30

Shifting time(sec)
Downshifting

6. Conclusion
With the aim of improving the rate of power
consumption of one-motor-type electric vehicles, aone-motor EV drive system was developed thatcombines a reducer with a two-speed gear shiftingmechanism and a drive motor.
Using a roller clutch in the shifting mechanism
enabled the reducer size and weight to be decreased.
At the same time, it was confirmed that shifting within
the required time of 0.3 sec was possible at speeds of
up to 80 km/h in a bench test and at speeds of up to
65 km/h in actual vehicles.
This potential for reducing the shifting time is great
due to possible measures such as improvement of
actuator performance and optimization of the control
method. NTN will make efforts to further improve theshifting performance and will continue practicalassessment tests on experimental vehicles.

Like the results of the bench tests, the tests on an
actual vehicle showed that the effect of the vehicle
speed on the shifting time is small. Unlike the benchtest, however, the shifting times for upshifting anddownshifting are about the same. This is due to theclutch disengagement waiting time, which wasconsidered only for downshifting in the bench test, but
was set also for upshifting in this test. This was donewith the aim of assuring the engagement anddisengagement of the clutch in actual vehicle running
conditions in which vehicle speed and torque varyfrom time to time.
In addition, a comparison between Figs. 6 and 12
reveal that, in spite of the same clutch disengagement
waiting time and the same vehicle speed condition,
the downshifting time on the test vehicle is longer by
0.02 sec than the result of the bench tests. The cause
for this time difference is thought to be due to the factthat the variation in the vehicle speed, namely theoutput rotation speed, in actual vehicle running causes
the synchronization time of the outer and the innerrings to increase. Furthermore, it can be supposedthat a greater variation of the difference in therotational speeds of the outer and the inner rings
occurs until the roller clutch closes after thecompletion of synchronization compared to the bench
test.

Reference
1) Eberleh, B. and Hartkopf, T., A high speed induction
machine with two-speed transmission as a drive forelectric vehicles, 2006 International Symposium on
Power Electronics, Electrical Drives, Automation andMotion, pp. 249-254, 2006.
2) Sorniotti, A., Boscolo, M., Turner A. and Caballino, C.,
Optimisation of a 2-speed gearbox for an electricvehicle, AVEC 2010, pp. 755-760, 2010.
3) Meier, T., Rinderknecht, S. and Fietzek, R., Electric
power train configurations with appropriate
transmission systems, 2011 SAE World Congress, No.
2011-01-0942, 2011.

Photo of authors

Fumihiro ISOBE

Yoshinori ITAKUTRA

Yusuke OSUMI

Guodong LI

EV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering

EV System Division
Driveline System
Engineering Dept.

EV System Division
Control System
Engineering Dept.

-41-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Paper ]Development of a New Steer-by-wire System

Katsutoshi MOGI*Tomohiro SUGAI*Ryo SAKURAI*
Nobuyuki SUZUKI*NTN has been developing a new steer-by-wire system. In addition to steeringfunctions, this steering system makes it possible to adjust the toe angle. The steeringsystem that was developed has fail-safe functions that can respond to various
failures in the system. We replaced the original steering system of an electric vehicle
with the one developed and evaluated the adjustable toe angle mechanism and thefail-safe function when a motor for steering is out of order. An analysis of vehicle
dynamics shows that the adjustable toe angle prevents wheels from slipping.

For this backup system, rather than using amechanical clutch plate, a system using
electromechanical features is used. Type III boasts
higher degrees of control and design, and canaccommodate novel car steering systems thatincorporate control means such as joysticks.
Aware of the greater freedom of Type III SBW
steering systems, NTN has developed a unique SBW(hereafter referred to as the NTN-SBW) throughimprovements to increase functionality while assuring
a high level of reliability. This paper reports on the

1. Introduction
Steer-by-wire (SBW) systems allow the amount ofsteering wheel operation to be transmitted in the formof electric signals to the vehicle wheels. Thesesystems help improve control performance for vehicle
safety while increasing vehicle design freedom. Thus,
this type of system seems to have promise as a nextgeneration automotive steering system. SBW systems
can be classified into the three types shown in Fig. 1.1)
The features of each type can be summarized asfollows.

1) Type I
A scheme in where the difference in the angle
between the steering wheel and steered wheels canbe controlled. This scheme is capable of controllingthe steering angle independently of the steered wheel
angle. However, since the steering reaction force istransmitted mechanically, this system cannot directly
control the steering reaction force.

Steering wheel
Differential angle
control function

Power steering

Steered wheel

2) Type II

Type I

When the system is in a normal state, there is no
mechanical link between the steering wheel and thesteered wheels. Note, however, this system isprovided with a backup feature that allows the steering
wheel to be linked to the steered wheels by means ofa clutch if the system fails. This system is also
capable of controlling steering reaction force.

Steering reaction
force actuator

ClutchSteering actuator

3) Type III
Type II

With this type, the system layout is the same as withthe type II system, but the backup system is different.

Type III

Fig. 1 Types of steer-by-wire systems*EV System Division Chassis System Engineering Dept-42-

Development of a New Steer-by-wire System

normally, the sub motor adjusts the toe angle of thesteered wheel, and if the main motor fails, the sub
motor functions as a backup motor for steering. The
toe angle adjustment function can contribute toincreasing the precision of vehicle control. 2) For thisreason, NTN’s new steering actuator can be regarded
as a mechanism that has achieved functional
improvement while assuring reliability.

results of a function assessment test of the NTN-SBW
and on theoretical consideration of the effectivenessof the toe angle adjustment function, which is afeature of the NTN-SBW.

2. NTN-SBW System
The NTN-SBW system consists of a steering
actuator, a steering reaction force actuator and acontroller.
The steering actuator drives a steering rod (a rackshaft) axially to steer the right and left wheels.
The steering reaction force actuator detects thesteering wheel angle, and simultaneously transmits
the state of the vehicle, which is conveyed from thetires, to the driver in the form of reaction force torque.
The controller is a unit that controls the steering
actuator and the steering reaction force actuator inharmony with each other and judges whether or not
the system is functioning normally to help the fail-safe
mechanism operate.

Fig. 2 Appearance of NTN's steering actuator

Gearing(for toe angle adjustment)

Primary motor(hollow)

Gearing(for steering)

Screw

Steering

2.1 Steering actuator
Fig. 2 shows the appearance of NTN’s newlydeveloped steering actuator, Fig. 3 depicts its
structure, and Table 1 summarizes its major
specifications.
As shown in Fig. 3, the steering actuator includes
two motors (main and sub), screws, a ball screw and aball spline. The ball screw shaft and the ball spline
shaft are arranged coaxially and coupled with eachother by means of screws, forming a steering rod.
The rotation of the main motor is converted intolinear motion via gearing by means of the ball screw.
This mechanism allows the steering rod to movelaterally, with the link mechanism consisting of a tie
rod and a knuckle arm to steer the tires. The
components shown in blue in Fig. 3 are those
associated with steering.
The rotation of the sub motor causes the ball spline
shaft to rotate via gearing, causing the coupling lengthof the screw to change. As the coupling length of thescrew changes, so does the total length of thesteering rod, allowing the toe angle of the tire to be
adjusted. The components shown in red in Fig. 3 areassociated with the toe angle adjustment. As shown in
Fig. 4, the toe angle refers to the angle of each of theright-hand and left-hand tires relative to the vehicle
traveling directly forward or backward as viewed fromabove.
As described later, should the main motor fail, thepower transmission path is switched over to the sub
motor, helping the vehicle maintain steering capability.
In this way, when the main motor is operating

Toe angle
adjustment

Secondary motor(hollow)#8
Ball spline

Ball screw

Steering rod

Fig. 3 Structure of steering actuator

Table 1 Specifications of steering actuator
Motor typeDC brushless

Steering system

Ball screw type

Power supply

48 V

Maximum thrust of steering rod

10 kN

Steering angle range when installed in vehicle±31 deg

Toe angle adjustment range when installed in vehicle±2 degTraveling
directionToe-inToe angle

Toe-out

Fig. 4 Definition of toe angle-43-

NTN TECHNICAL REVIEW No.79（2011）

2.2 Steering reaction force actuator

actuator system to make adjustments to obtain alaterally-equal toe angle. It is possible to set up a
steering angle on either side independently of theother to cope with the current traveling conditions of
the vehicle.

Fig. 5 shows the appearance of NTN’s new steering
reaction force actuator, Fig. 6 depicts its structure,
and Table 2 summarizes its major specifications.
The torque damper, the steering angle limiting
mechanism, the reducer, the motor and the angle
sensor are installed on a common shaft to reduce thesize and weight of the actuator.
In order to provide the driver with a comfortablefeeling when steering, the damping effectiveness ofthe torque damper is increased by filling the space
between the output shaft and housing with a highviscosity fluid.

Steering wheel

2.3 Controller
Fig. 7 is a block diagram of the control system of
NTN’s new SBW system. The electronic control unit
(ECU) consists of the three ECUs—ECU-A, ECU-B
and ECU-C—that control the actuator motors. The
differential torque between the steering torque Th ofthe driver and the steering reaction force torquecauses the steering wheel to rotate, and the angle of
the rotation is detected by the angle sensor. According
to the steering wheel angle, ECU-A controls the
position of the steering mechanism.
The external force acting on the steering
mechanism is detected by a steering torque estimator.
Then, a target reaction force torque is generated on
the basis of the external force detected, and according
to which the ECU-C controls the torque of the reaction
force motor.
The vehicle control unit (VCU) that controls the
motion of the entire vehicle sends toe angle commandsignals to ECU-A and ECU-B to control the main and
sub motors of the steering actuator in coordination.
This makes the center line of the steering column
coincide with the lateral midpoint of the steering

Fig. 5 Appearance of reaction force actuator

Motor

ReducerAngle sensor

Table 2 Specifications of reaction force actuator
Motor type
Power supply

48 V
15.5 Nm

Steering wheel angle resolution

0.01 deg

ECU-B＋ −Angle/position−

Steering
wheel

Angle
sensor＋DC brushless

Maximum reaction force torqueθt＋Torque damper

Fig. 6 Structure of reaction force actuator

VCU

Th

Steering wheel angle
limiting mechanism

θs
ECU-A

Secondarymotor

Toe angle adjustment
mechanism

Position sensor

Primary
motor

Steering mechanism−

Reactionforce motor

ECU-C

Angle/position

Position sensor＋ Target reaction force
torque generator

Fd Steering torqueestimator−
Torque sensor

Fig. 7 Block diagram of control system-44-

Development of a New Steer-by-wire System

3. Fail-Safe Mechanism

4. Verification of Motor Switching
Mechanism

The most critical challenge for an SBW system is tomaintain its functionality even when there is a failure,in other words, to ensure its fault tolerance. Thissystem was designed by taking into account the loseof power, disconnected wires, failed motors or
sensors, malfunction and failure of the ECU, andfailed mechanisms. This section describes the motor
switching mechanism of the steering actuator thatswitches over to the sub motor to maintain steering
capability even if the main motor has failed for any
reason.
Fig. 8 schematically illustrates the state of asteering actuator with a failed main motor. The
schematic diagram of the steering actuator in Fig. 3
shows the case when the main motor condition isnormal.
At the end of the motor switching shaft thatpenetrates the main motor, a spring is provided toswitch the power transmission mechanism. When themain motor operates normally, this spring ismaintained in an energized state. Should the main
motor fail, the spring is released, with the motor
switching shaft moved to the left by the compressive
force of the spring as shown in Fig. 8. Consequently,
the clutch mechanism installed in the motor switching
shaft trips to create a new power transmission pathfrom the sub motor to the ball screw. At the sametime, the power transmission from the sub motor to theball spline is interrupted. Even if the rotor of the main
motor seizes, the power is transmitted because themotor switching shaft penetrates the hollow rotor. In
this way, the sub motor helps maintain the wheel
steering function of the SBW even when the main
motor has failed.

4.1 Bench assessment test
To verify operation of the motor switchingmechanism of the steering actuator, a bench test was
conducted. Fig. 9 illustrates the configuration of thetest rig used for this bench test.
Assuming that the main motor failed while thevehicle was turning with the steering wheel maintained
at a certain position, the steering rod of the steering
actuator was given three levels of axial load—0 kN, 3
kN and 6 kN. The load was applied by an air cylinder.
Because of the assumption that the steering wheel
was kept at a particular angle, the axial position of thesteering rod was kept controlled.
Under these conditions, the switching time andquality of the motor switching mechanism wereassessed. The switching time was defined as the timespan required, beginning with sending the commandsignal for releasing the spring and ending withcompletion of the movement of the motor switching
shaft to the specified position. The switching was
judged to have been normally performed if, after thecompletion of the switching, problem-free lateral motionof the steering rod by the sub motor was confirmed.
Fig. 10 graphically illustrates the motor switching
test. To determine switching reliability, three test runs
were performed for each loading condition.

Linking bar

Air cylinder

Load cell
Load

Load

Steering actuator
Primary motor(Failed)Motor switching
shaft

Load cell

SpringAir cylinder
Linking bar
Secondary motor

Wheel
steering

Fig. 9 Bench assessment test rig

Switching time (msec)

20

Fig. 8 Schematic view of malfunctioning steering actuator

15
0kN
3kN

10

6kN
5

0

1st

2nd

3rd

Fig. 10 Time required to complete motor switching-45-

NTN TECHNICAL REVIEW No.79（2011）

apparent that the time required for motor switching is
10 msec, which is about the same as the bench test
result, and that the wheels could be steered normallyafter motor switching. The signs of disturbance in thesteering before and after switching were due to thedriver’s steering action, rather than a malfunction of
the switching mechanism.

From Fig. 10, the switching time falls in a range of
10–13 msec and normal switching occurred
regardless of the load. As the load became larger, themotor switching action became slightly slower. The
reason for this is that an increase in the torque appliedon the key and spline on the motor switching shaft
causes the slide resistance at the time of switching toincrease.
For this test, the time needed to judge whether themain motor was normal was ignored. Even when thistime is taken into account, the time elapsed from thefailure of the main motor to the completion of theswitching was 0.1 sec or less.

CommandsignalMotor switching shaft
position108

Switchingtime

4.2 Onboard assessment test
Next, we assessed the operation of the motor
switching feature of the NTN-SBW on a test vehicle
that had its EPS replaced with the NTN-SBW.
Table 3 shows the main specifications of the test
vehicle. Fig. 11 shows a view of the NTN-SBW
installed in the vehicle.
In the test, the vehicle was slalomed at 60 km/h,
and a motor failure signal was sent to the controller totrip the motor switching mechanism.
Fig. 12 shows the test results. The upper chartshows the main motor failure signal and the axial
position of the motor switching shaft, and the lower
chart shows the steering wheel angle and the yaw
angular velocity on the vehicle. From Fig. 12, it is

6
4
2

0.45

0.47

0.49

0.51

0.53

0
0.55

Shaft position (mm)Main motor failure

Time (sec)

Yaw angularvelocity
40

90

20

0

0

Steering wheel angle
(deg)

180

-20

-90

-180
0

Main motor failure
1

2

3

4

-40
5

Yaw angular velocity(deg/s)Steering wheel
angle

Time (sec)Fig. 12 Results of evaluation of motor switching systemin test vehicle

Table 3 Specifications of the test vehicle
Drive system

Front wheel drive

Total length

3765 mm

Total width

1690 mm

Total height

1510 mm

5. Verification of Toe Angle Adjustment
Mechanism
NTN's newly-developed steering actuator isequipped with a toe angle adjustment mechanism.
Controlling the toe angle in accordance with thetraveling state of the vehicle in cornering could help
prevent slipping of the vehicle by appropriately varying
the proportion of lateral force between the right andleft wheels. As one of the basic tests for this control
feature, a cornering test was conducted on the samevehicle as described previously.
In the test, five levels of toe angle—0 degs, ±1.0
deg and ± 1.8 degs—were applied to the front wheels.
The vehicle was allowed to circle at 40 km/h at aradius of 30 m, and the lateral force acting on the front
wheels was investigated. The toe angles applied to
the front wheels were estimated using calculations
based on the coordinate system of the link mechanism

Steering actuator

Steering reaction force actuator
Fig. 11 View of NTN-SBW mounted in test vehicle-46-

Development of a New Steer-by-wire System

consisting of the steering rod, tie rod and knuckle arm.
The lateral force was measured by means of a load
sensor built into the hub. For the toe angle, positivemeans toe-out and negative means toe-in.
Fig. 13 shows the lateral force on the right and left
wheels during cornering. With the vehicle cornering at
a constant speed, the lateral force is theoreticallyconstant. Two cornering directions were used—
clockwise (CW in Fig. 13) and counterclockwise(CCW in Fig. 13) with the vehicle as seen from above.
The lateral force acting leftward with respect to thetraveling direction was taken as positive.
As shown in Fig. 13, the inner lateral force on thetire increases (and the outer lateral force on the tiredecreases) in the toe-out condition, while the innerlateral force on the tire decreases (and the outer
lateral force on the tire increases) in the toe-in
condition. It was confirmed that adjusting the toe angle
causes the ratio of the lateral force on the left wheel tothe lateral force on the right wheel to change.

Left wheel (CCW)

Right wheel (CCW)

Left wheel (CW)

Right wheel (CW)

φ Cφ

ψ

Traveling
direction

Fig. 14 Vehicle motion model 4)(yaw) and around the x axis (roll).
･ The height of the roll center differs between the front
and the rear of the vehicle.
･ The model is based on a four-wheel vehicle in orderto take differences between the lateral force acting
on the right wheel and that acting on the left wheel
into consideration.
･ The lateral force on a tire corresponds with thedirection of the Y axis.
･ The vehicle advances in the X-axis direction(forward) at a constant speed V.

4000

Lateral force (N)

Kφ

3000
2000

6.1.1 Equation of motion

1000
0

On the basis of the above assumptions, theequation of motion with respect to the paralleladvancement in the Y-axis direction in this model isexpressed by the following equation:

-1000
-2000
-3000
-4000
-2

-1

0

1

2

Toe angle (deg)mV (β+ψ)＝FfL+FfR+FrL+FrR ……………（1）Fig. 13 Effect of toe angle on lateral tire forces

In equation (1), m denotes the vehicle mass and F
is the lateral force acting on individual tires. In thesuffixes to lateral force F, the first letter refers to thefront or rear and the second to the left or right. For
example, FfL refers to the lateral force acting on thefront left wheel.
In the similar way, the equation of motion around
the Z axis (yaw) is given by:

6. Theoretical Consideration
6.1 Vehicle motion model
As described earlier, it was confirmed in actual
vehicle testing that the toe angle affects the balance inthe lateral force in the steered wheel. In this section,
the effect of toe angle adjustment on the vehicle
motion is theoretically considered by means of avehicle motion model.
This vehicle motion model is formed on the basis of
Minakawa’s vehicle motion model 3) shown in Fig. 14,
with the following assumptions. 4), 5)･ The model is one with three degrees of freedom
describing the motion in the y axis, around the z axis

IZ ψ+IXZφ＝lf (FfL+FfR)-lr (FrL+FrR) ………（2）In the equation, IZ denotes the yaw moment ofinertia, IXZ the yaw-roll inertia product, If and Ir the
distance from the vehicle center of gravity to the front
axle and to the rear axle, respectively.
The equation of motion around the X axis (rolling) is
given by:-47-

NTN TECHNICAL REVIEW No.79（2011）

IXφ+Cφφ+(Kφ-mghc)φ+IXZψ
＝hf (FfL+FfR)+hr(FrL+FrR)

δis the steering angle of the front wheel, formed by
the orientation of the vehicle (X-axis) and the surfaceof revolution of the tire. δt is the toe angle of the front
wheel and, as in the previous chapter, toe-out is taken
as positive. Cr denotes the roll steer coefficient andCrr is the roll rate steer coefficient.

……………（3）In the equation, IX denotes the roll moment ofinertia, Cφ is the roll attenuation factor, and Kφ is theroll rigidity. The roll moment arm length at the vehicle
center of gravity hc is expressed as shown below,
using the roll moment arm lengths hf and hr at thefront and rear axles.
hc＝6.1.3 Tire characteristics
The cornering power Kc is expressed in thefollowing equations using the lateral rigidity of the tiresuspension KY and the steady cornering power K.

hf lr+hr lf………………………………（4）lf + lr

1
-Csff
Kf

（

By solving the three coupled equations of motion
above and determining the angle of sideslip of the
bodyβ, the yaw angular velocity ψ, and the roll angle
φ, the state of motion of a vehicle can be determined.

（）
（
（）
（

Kcf +

1
Kcr +
-Csfr
Kr

（

The lateral force to the tire expressed in equations(1) to (3) is expanded in the following equations:＝
FfL＝-KcfLβfL
FfR＝-KcfRβfR
FrL＝-KcrLβrL
FrR＝-KcrRβrR
In the above equations, signs such as βfL denotethe sideslip angle of each tire. As shown in Fig. 15,
the sideslip angle of a tire refers to the angle formedby the traveling direction and the surface of revolutionof the tire. Kc is the cornering power of the individual
tires. The sideslip angle of a tire is expanded further
by the following equations:l fψhfφ-δ-δt+-Crfφ-CrrfφVV

βfR＝β+

l fψ
hfφ-δ+δt+-Crfφ-Crrfφ …（6）VVβrR＝βrL＝β-

｛
｛

ΔWf＝

Kcf

-1

1
KYfV

-Csfrf…………（7）-1

1-Csfrr

KYrV

Kcr

-1

1
KYrV

-CsfrrKφφ+ Hf (FfL +FfR)
2

｝
｝

Kφ
φ+ Hr (FrL +FrR)
ΔWr＝
2

df
…………（8）dr

In the above equations, H denotes the height of theroll center of the front and the rear and d the track
width of the front and the rear. On the basis of this, thecornering power is used in the corrected forms as in
the following expressions. 6)lrψ hrφ+-Crrφ-Crrrφ
VV

Travel direction of the tire

）
）
）
）

In the above equations, Csf denotes the lateral force
steering coefficient and Csfr is the lateral force
differential steering coefficient. The cornering power
varies depending on the vertical force W acting on thetire. With rolling motion occurring on the vehicle, thevertical force acting on the tire increases or decreases
by ΔW as shown in the following equations:………………………………（5）βfL＝β+-Csfrf

KYfV＝

6.1.2 Tire lateral force model-1

1

KcfL ＝

WfL-ΔWf
WfL

KcfR ＝

WfR+ΔWf
WfR

（
（
（
（

Vehicle orientation

Tire orientation

WrL-ΔWr
KcrL ＝WrL

Steering angle plus
toe angle

KcrR ＝Sideslip angle

Fig. 15 Tire slip angle and steering angle

-48-

2/3

）
）
）
）

Kcf

2/3

Kcf

2/3

Kcr

WrR+ΔWr 2/3
Kcr
WrR…………………（9）Development of a New Steer-by-wire System

6.2 Conditions for analysis

operated to provide a 1-deg. toe-in setting, the F/W on
both sides becomes smaller than 0.7, preventing thefront wheels from slipping. Furthermore, with a toe-in
setting at the cross-section of the two straight lines,
namely, at about 0.5 deg., the F/W values on bothsides are equal, decreasing the risk of the front wheellosing traction. The optimum toe angle that equalizes
the F/W values on both sides has a relationship withthe vehicle speed as shown in Fig. 17. Using the
NTN-SBW allows one to adjust to the optimum toe
angle while running.
Provided with the characteristics described above,
the toe angle adjustment mechanism, a feature of the
product developed, is capable of providing an
optimum toe angle that allows a vehicle to run safely,
avoiding the danger of slipping even on road surface
conditions that invite tire slipping.
Adjusting the toe angle to about zero degrees forstraight-ahead driving allows the lateral force on thetire to decrease, with an improvement in fuel efficiency
to be expected.

Table 4 shows the main conditions. The main
parameters in the table are values for theexperimental vehicle determined on the basis of theresults of the running test. For the experimental
vehicle used this time, the effect of the inertia product
IXZ was so small that it was approximated as zero.
In the analysis, the steering angle was set as aconstant and cornering on a static circle wasassumed.

Table 4 Major conditions for analysis
Vehicle mass m

1400 kg

Distance from vehicle center of gravity to front wheel lf

1.1 m

Distance from vehicle center of gravity to rear wheel lr

1.29 mFront wheel track width df

1.46 mRear wheel track width dr

1.47 mRoll inertia moment IX

600 kg・m2

Yaw inertia moment IZ

1900 kg・m2

Roll rigidity Kφ

5.0×104 Nm/rad

Roll attenuation coefficient Cφ

5500 Nm・s/rad

Front wheel static cornering power Kf

4.3×104 N/rad

Rear wheel static cornering power Kr

3.8×104 N/rad

Tire suspension lateral rigidity KY

2.0×105 N/m

1.4Dangerrange

1.2

Front left wheel

F/W

1.0

6.3 Results of analysis
The results of the analysis were evaluated by theratio of the resultant braking/driving force Q and thelateral force F to the tire vertical force W.
With the running speed taken as a constant in thisanalysis, the braking/driving force Q was ignored, withthe result of the analysis evaluated by the ratio of F to
W as shown in equation (10). This value does notexceed the coefficient of road surface frictionμ. If it
did, the tire would slip.
F
W

0.8
0.6
0.4
0.2
0
-2

-1

0

1

2

Toe angle (deg)Fig. 16 Relationship between toe angle and F/W values

0.10………………………（10）Optimum toe angle (deg)Q2 + F 2
W

Front right wheel

Fig. 16 shows some of the results of analysis. This
analysis is based on a hypothetical situation in which
a vehicle corners at a speed of 60 km/h on a circle
with a radius of 50 m. The direction of cornering iscounterclockwise based on the vehicle seen fromabove.
Fig. 16 shows that, with the coefficient of roadsurface friction μ assumed to be 0.7, a toe-out setting
causes the front left wheel to slip. A 2-deg. toe-in
setting also causes the front right wheel to slip. If thefront right wheel, which needs large lateral force slips,
the driver becomes unable to control the vehicle,
resulting in a very dangerous condition.
However, if the toe angle adjustment mechanism is

-0.1
-0.2
-0.3
-0.4
-0.5
-0.6
10

20

30

40

50

Vehicle speed (km/h)

Fig. 17 Optimum toe angles for vehicle speeds

-49-

60

NTN TECHNICAL REVIEW No.79（2011）

7. Conclusion
Reference

The NTN-SBW is equipped with a fail-safe
mechanism that addresses failures of the system as awhole. As an example, a fail-safe demonstration test
was conducted on the assumption that the main motor
for steering failed. In addition, a bench test and an
actual-vehicle test were conducted to demonstrate theeffectiveness of the toe angle adjustment function.
Furthermore, a vehicle motion model was created totheoretically assess the relationship between the toeangle and tire slip. The results obtained aresummarized as follows:･ It was demonstrated in a fail-safe experiment thatswitching from the main motor for steering to thesub motor does not hinder continued steering
capability.
･ A running test on an actual vehicle showed that
adjustments to the toe angle cause the balance inlateral force between the left and the right tires tochange.
･ A numerical analysis showed that even in a running
condition in which the tire would lose traction, a toe
angle adjustment could make it possible to avoid theloss of tire traction.

1) Motoyama, S., Possibilities of Steer-by-wire on
Vehicle Dynamics, Journal of Society of Automotive
Engineers of Japan, Vol. 57, No. 2, pp. 39-43, 2003.
2) Yamamoto, M., Effect of Wheel Alignment on
Handling and Stability, Journal of Society ofAutomotive Engineers of Japan, Vol. 54, No. 11, pp.
10-15, 2000.
3) Minakawa, M., Model Concept for a Simple 3DOF
Vehicle Model with Considerations of the Effect ofBody Roll, Transactions of the Society ofAutomotive Engineers of Japan, 20075786, 2007.
4) Abe, M., Vehicle Dynamics and Control, First
edition, Tokyo Denki University Press, 5-180, 2008.
5) Sakai, H., Influence of Roll Characteristics on
Vehicle Dynamic Behavior, Transactions of theSociety of Automotive Engineers of Japan,
20055693, 2005.
6) Kobayashi, A., The Latest Automotive Engineering,
Tosho Shuppan, 1975.

Photo of authors

Katsutoshi MOGITomohiro SUGAI

Ryo SAKURAI

Nobuyuki SUZUKIEV System Division
Chassis System
Engineering DeptEV System Division
Chassis System
Engineering DeptEV System Division
Chassis System
Engineering DeptEV System Division
Chassis System
Engineering Dept

-50-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Article ]Development of Two-seat Electric Vehicles
with In-Wheel Motors

Akira YAMAGATA*
Chinami ITOU*
Kayo SAKAI*
Kaori TOMODA*

Aiko ISHIKAWA*
Tomomi GOTOU*
Yukiko TAKEKAWA*
Daisuke MATSUOKA**

Compact electric vehicles are thought to be suitable as small models due to theadvantages of quietness, low-heat generation and no exhaust gas. In particular, in-wheel
motors realize smaller space requirements for power trains and drive trains. Moreover,in-wheel motors can achieve high-performance drive through independent steering
mechanisms for all wheels. An NTN development team of young engineers has designed
a two-seat electric vehicle that provides “pivot turning” and “lateral movement”
capabilities with in-wheel motors. This paper introduces the features and structure of thiselectric vehicle.

example, the development of a city-commuter-typevehicle on which both the track width and the wheelbase are variable as a result of adopting a four-wheeldriven, two-tier, four-wheel independent steering
mechanism has been reported. 5)
Having been engaged in the development ofsystems for EV use such as in-wheel motor drive
systems 6) and steer-by-wire systems, 7) NTN has
realized vehicles equipped with its EV-specificsystems to demonstrate the features and advantagesof such newly-developed systems.
Given these conditions, NTN has organized adevelopment team consisting of young engineers toinvestigate, the uses for small EVs equipped with inwheel-motor-driven systems. As a result, they haveinvented the "Electric Mobility," a small two-seatvehicle that makes the most of the features of inwheel motors, is capable of four-wheel independentwide-angle steering, and is compact and excellent insmall-radius turn performance.
This paper presents the concept of NTN's newlydeveloped two-seat Electric Mobility and the featuresof its mechanisms.

1. Introduction
With automobile manufacturers having started
mass-producing passenger car-type electric vehicles(EVs) recently, EVs are seeing rapid popularization.
On the other hand, some automobile manufacturers
have proposed small EVs with one or two seats andelectric two-wheel vehicles that are intended for shortdistance travel, 1)-4) and some have already beenbrought to market.
Electrification is suitable for small vehicles aselectric power is quiet, generates little heat, and has
no exhaust gasses when compared to similar vehicles
with internal combustion engines. Such compact
vehicles are expected to be used as means oftransportation not only on public roads but also inindoor spaces, such as hospitals and factories. In
particular, the application of in-wheel motors as powersources offers the possibility of two great advantages.
One of the advantages is freedom from the necessityof installing differentials and constant velocity joints(CVJ) for the drive wheels, so space is easily madebetween the drive wheels. The other is the wide
steering angle range, which is not limited by CVJ
operating angles.
Several EVs that are equipped with in-wheel motors
and said to be capable of four-wheel independentwide-angle steering have been proposed thus far. For**EV System Division Driveline System Engineering Dept.
**EV System Division Control System Engineering Dept.

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NTN TECHNICAL REVIEW No.79（2011）

wheel-drive, four-wheel-drive or even one-wheel-drive
mode. The steering angle required to achieve a pivotturn is determined by the wheelbase and the track
width. For example, when the wheelbase equals the
track width, the steering angle required is 45 degrees.
When the wheelbase is longer than the track width asin common passenger automobiles, a steering angleof between 45 degrees and 90 degrees is required.
For lateral movement, the four wheels are all turned to
90-degree angles.
Table 1 shows the rotation directions of individual
wheels in different running modes.
The rotation directions of individual wheels during
pivot-turning and lateral movement differ from thoseduring normal running.
These movements cannot be achieved in ordinary
four-wheeled vehicles equipped with differentials.
They are achieved only by the combination of the inwheel-motor drive system, which is characterized byindependent drive for the right and left wheels, and amechanism that steers the four wheels in wide angles
independently.

2. Vehicle Concept
The considerations of the development team have
led to the conclusion that for the popularization of EVs
it is important that NTN study the commercial
usefulness of small vehicles. These vehicles excel inhandling in an urban environment due to theircompact bodies and feature sets that contain certain
levels of road running performance and load-bearing
capacity.
The two-seat Electric Mobility vehicle presented inthis paper has a compact body that excels in thecapability of turning in a small radius in a limited
space. Furthermore, an advantage of in-wheel-motorequipped vehicles is that the four-wheel independentwide-angle steering feature makes it possible to pivotturn and to move laterally. Such compact electricvehicles will work not only as means of everyday
transportation, they also have the potential to be usedin other practical applications, such as medical
services and tourism.
Fig. 1 shows the orientations of different wheels
during pivot turns and lateral movements. A pivot turnis a rotational movement of a vehicle with the vehicle
center as the center of rotation. The “vehicle center”
here refers to the point of intersection of the bisector
between the front and rear wheel axes and of thecenter line between the right and left wheels. When avehicle pivot-turns, the number of drive wheels does
not affect whether or not it is possible. In other words,
if the rotation axes of all wheels are directed toward
the turning center of the vehicle, a pivot-turn ispossible regardless of whether the vehicle is in two-

Table 1 Rotational direction of wheels
Running mode
Normal running(forward)
Pivot-turning(right)Lateral movement(rightward)

Front left Front right
wheel
wheel

Rear left
wheel

Rear right
wheel

CCW

CW

CCW

CW

CCW

CCW

CCW

CCW

CCW

CCW

CW

CW

CW: Clockwise as seen from the outside of the vehicle
CCW: Counterclockwise as seen from the outside of the vehicle
Note: Rotation is reversed for backing up, left pivot turns and left lateral
movements

Front left wheel(a) Normal running(b) Pivot turn
：Steered turning center

Fig. 1 Pivot turn and lateral movement-52-(c) Lateral movement

Development of Two-seat Electric Vehicles with In-Wheel Motors

steering gearbox is transmitted via a tie rod to the hub
carrier to produce steering torque around the kingpin
shaft. In this test vehicle, however, the steering range
is wide, and moreover, it is necessary for the right andleft wheels to be steered independently of each other.
For this reason, a steering system with a tie rod
scheme results in a very long operating stroke in thesteering gearbox, requiring a large area. In this test
vehicle, therefore, the four wheels are each equippedwith a steering actuator with the aim of minimizing thearea occupied by the steering system and thereby
making full use of the space around each wheel.
Fig. 3 shows the layout of an in-wheel motor andsteering actuator. The suspension is structured on thebasis of a four-wheel double wishbone system.
The in-wheel motors and the steering actuators arebuilt into an integral structure, and the rotation aroundthe kingpin shafts of the output members of thesteering actuators connected to the upper arm joints isregulated. Thereby, this structural arrangement allows
the upper arms to support the steering reaction force.
This arrangement allows the steering actuator bodies
to rotate around the kingpin shafts, allowing the inwheel motors connected to the steering actuators andthe wheels connected to the in-wheel motors to besteered.

3. Test Vehicle
3.1 Vehicle layout
Fig. 2 shows the test vehicle. The vehicle isequipped with two rear wheels that have in-wheel
motors. To enable four-wheel independent steering,
all four wheels are equipped with steering actuators.
Furthermore, in order to reduce space, the steering
actuators are located near the wheels.
The major dimensions of the body are about 1700
mm in total length, about 1200 mm in total width andabout 1900 mm in total height. Including space
sufficient for two passenger seats, an inverter, abattery and other components, the vehicle wasdesigned for maximum compactness. Since thebattery is the heaviest component, it was installed in
the lowest part at the center of the body, with thespace above the battery serving as the passenger leg
space.
Furthermore, by providing the seating space above
the rear wheels and installing the inverter above thefront wheels, the space needed for this vehicle is as
small as the area occupied by two ordinary mopeds.

3.2 Suspension
The steering mechanism for conventional vehicles
is a tie rod system in which the thrust from theUpper arm joint
Inverter

Tire/wheel
Upper arm

Steering actuator

Kingpin shaft

In-wheel motor

Lower arm

Battery
Rear wheel(equipped with in-wheel motor)

Steering
actuator

Lower arm joint

Fig. 2 Test vehicle

Fig. 3 Suspension layout-53-

NTN TECHNICAL REVIEW No.79（2011）

3.3 Kingpin offset

3.4 In-wheel motor unit

Fig. 4 schematically illustrates the kingpin offset.
The kingpin offset refers to the distance between theintersection of the kingpin axis with the ground and thecenter of the tire tread. It is general practice tominimize the kingpin offset with the steering wheel.
The offset is determined through comprehensive
considerations around possibly conflicting factors suchas straight running stability and minimum necessary
steering wheel operation effort. To reduce the kingpin
offset, the joint of the lower arm must be placed inside
the wheel.
However, because the range of the steering angle
on this vehicle exceeds 120 degrees, the location of
the lower arm joint is greatly limited. For example,
when the clearance between the outer circumferenceof the motor and the wheel bore surface is small andthe lower arm joint is located in the bottom of thewheel in order to minimize the kingpin offset, if thesteering handle is turned greatly, then the lower arm
can interfere with the inside surface of the wheel.
In this test vehicle, the lower arm joint is situatedoutside the wheel (inside the vehicle) so that the lower
arm does not interfere with the wheel when thesteering wheel is turned at a large angle. The resulting
greater kingpin offset can be one factor that
contributes to an increase in steering torque. Thiseffect is particularly apparent when the steering wheel
is turned sharply while the vehicle is stationary andthe brake is applied. For this reason, the tires on thisvehicle are set free with the brakes not applied whensteering in a wide angle to avoid an increase in thesteering torque.

An in-wheel motor composed of a motor, a reducer
and a hub is used for the drive wheels. With thecondition that it is used with a wheel of 14-inch
diameter or greater, this in-wheel motor was designedwith emphasis placed on a greater reduction of thedimension in the axial direction than in the motor
outside diameter.

3.5 Steering actuator
Fig. 5 shows the structure of the steering actuator.
In the steering actuator, the motor output speed isreduced by a worm gear to produce steering torque.
With the shaft of the worm wheel forming part of the
kingpin shaft, the worm wheel is fastened to the upperarm so that only the upper arm joint is free to pivot.
This configuration restrains the rotation of the worm
wheel. For this reason, as the motor built into thesteering actuator rotates, the housing covering thepart from the worm gear to the motor and the coupledin-wheel motor rotate around the kingpin shaft.

Upper arm joint

Upper arm

Kingpin shaft
Worm wheel

Worm
Driven gear

Housing

Drive gear
(motor output)

Fig. 5 Structure of the steering actuator

Kingpin offset(a) General layout(b) Developed vehicle：Joint position

Fig. 4 Kingpin offset and joint position-54-

Development of Two-seat Electric Vehicles with In-Wheel Motors

4.2 Steering control

4. Mode Operation Control

The steering system of this vehicle is provided by asteer-by-wire scheme, with the individual wheels being
capable of independent steering. In the normal
running mode, however, only the front wheels aresteered, as with conventional vehicles, according to
the steering input. However, it is possible to reduce
the turning radius by steering the front and the rear
wheels in mutually opposite phases. In addition, in asituation such as a lane change, it is possible toreduce the yaw of the vehicle and improve roadholding by steering the front and the rear wheels in the
direction of the same phase.
In contrast, the pivot-turning mode enables all four
wheels to be steered so that their axes are oriented
toward the turning center of the vehicle. In lateral
movement mode, the steering angle is increased
beyond that used in pivot-turning mode, with all four
wheels turned a full 90 degrees from normal mode.
Figs. 7 and 8 show how the rear wheels of the test
vehicle are steered in pivot-turning mode and lateral
movement mode, respectively. The front wheels canbe steered in the same way, and it is even possible for
all four wheels to be turned by over 90 degrees within
a limited vehicle space.

This test vehicle is equipped with steering
mechanisms that are mechanically independent on
the four wheels, and the four wheels are steered byelectric signals transmitted through the steer-by-wire
control system.
Fig. 6 shows the basic control configuration inwhich the in-wheel motor is installed in all four wheels.
The vehicle is integrally controlled by the vehicle
control unit (VCU), which provides overall control of thein-wheel motors and the steering actuators according
to the three running modes---normal running mode,
pivot-turning mode, and lateral movement mode.
According to the running mode selected by theoperator and the running commands, including theaccelerator stroke, the VCU gives commands for thedrive power and the rotational direction needed forindividual in-wheel motors to the inverter. Based onthese commands, the inverter generates thenecessary AC outputs. At the same time, in responseto the steering commands from the VCU, the steering
drivers in turn control the steering actuators so thatthe target steering angle is achieved.

4.1 Control of drive wheels
When the vehicle advances (or backs up) in normal
running mode, the drive wheels on both sides drive
the vehicle forward (or backward) in a manner
identical to that of the drive wheels of conventional
vehicles. On the other hand, in pivot-turning mode andlateral movement mode, the drive wheels move in the
directions necessary for these modes as shown in
Table 1.

In-wheel motors
Steering actuators

Fig. 7 Steering in pivot-turning mode
Main battery

Steering
drivers

InverterPower line
Signal line

VCU
Steering actuators
In-wheel motors

Fig. 6 Control system configuration diagram

Fig. 8 Steering in lateral movement mode

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NTN TECHNICAL REVIEW No.79（2011）

5. Two-seat Electric Mobility

6. Conclusion

Fig. 9 shows an example of a vehicle design
embodying the vehicle concept described in thispaper, and Table 2 lists the major specifications forthis design. This vehicle is equipped with in-wheel
motors installed in all four wheels, is four-wheeldriven, and is steered using the four-wheel
independent wide-angle steering principle. With alayout following that of the test vehicle, this vehicle
was designed so as to locate the battery and theinverter below the floor to lower the center of gravity ofthe vehicle with the aim of improving the load-capacity
and securing space for two seats.

A two-seat electric vehicle with a compact body was
fabricated that uses a four-wheel, wide-angle steering
mechanism and in-wheel motors, making it capable ofboth pivot-turning and lateral movement.
Equipping each wheel with a small-footprint steering
actuator makes it possible to turn all four wheels over
90 degrees in a limited vehicle space.
In the future, this test vehicle will be used to confirm
the function of EV systems.

Reference
1) Society of Automotive Engineers of Japan, Hybrid
Vehicles, Fuel Cell Electric Vehicles, Electric Vehicles,
Journal of Society of Automotive Engineers of Japan,
Vol. 61, No. 8, pp. 84-89, 2007.
2) Nunogaki, N. et al., Personal Mobility Design ofToyota, Journal of Society of Automotive Engineers ofJapan, Vol. 64, No. 6, pp. 56-59, 2010.
3) Makiyama, T. et al., 41st Tokyo Motor Show: Nissan
leads the way to zero-emission, Nissan Technical
Review, No. 66, pp. 73-81, 2010.
4) Nishiyama, M., Development of the electric commutermodel EC-03, 1st International Electric Vehicle
Technology Conference 2011 (EVTeC'11), No.
20117222, 2011.
5) Fukui, A. et al., 40th Tokyo Motor Show, Nissan
Technical Review, No. 62, pp. 66-75, 2008.
6) Suzuki, M. et al., Development of an In-Wheel Motor
Axle Unit, NTN Technical Review, 75, pp. 46-52, 2007.
7) Mogi, K. et al., Development of Steer-by-wire Systemwith an Adjustable Toe Angle Mechanism,
Transactions of the Society of Automotive Engineers ofJapan, No. 146-11-257, 2011.

Fig. 9 Two-seat Electric Mobility

Table 2 Main specifications of the vehicle
Battery

In-wheel motors
Steering system
Cooling system
Seating capacity

DC 96 V(Serving both the in-wheel motors
and the steering actuator)
Maximum output/shaft

2 kW

Four-wheel drive
Drive wheel
Four-wheel independent wide-angle
steering mechanism
In-wheel motor

Air-cooled

Inverter

Water-cooled
Two people-56-

Development of Two-seat Electric Vehicles with In-Wheel Motors

Photo of authors

Akira YAMAGATA

Aiko ISHIKAWA

Chinami ITOU

Tomomi GOTOU

Kayo SAKAIEV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

Yukiko TAKEKAWA

Kaori TOMODA

Daisuke MATSUOKAEV System Division
Driveline System
Engineering Dept.

EV System Division
Driveline System
Engineering Dept.

EV System Division
Control System
Engineering Dept.

-57-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Hub Bearing with Integrated Multi-axis Load Sensor

Kentaro NISHIKAWA*

Cars have advanced vehicle control systems to ensure safety and various
sensors installed to sense and control vehicle conditions. However, thecontrol responses of conventional yaw and G sensors installed in vehicles
are delayed because they begin to detect after the vehicle position changes.
To further improve vehicle control, a hub bearing with an integrated multiaxis load sensor has been developed that can detect the frictional forcechange of a tire at a position near the road surface in real time.

1. Introduction

2. Development Concept and Structure

To ensure safety, recent automobiles have highlysophisticated vehicle control systems that include
various onboard sensors for helping estimate andcontrol the current states of the vehicle. A typicalcontrol system for this purpose receives signals fromthe yaw rate sensors and accelerometers installed in
the car body (onboard sensors), and from velocity
sensors (ABS sensors) and steering angle sensors on
the axle side; on the basis of information from thesesensors, the control system estimates the current
state of the vehicle. Note that the onboard sensors
begin detecting changes in the vehicle after a change
in the vehicle dynamics has occurred; consequently, avehicle control sequence begins after a certain delay1).
For further sophistication of vehicle control systems,
NTN believes that frictional force between tire tread
and road surface should be detected on a real timebasis at a location nearest to the road surface in orderto promptly determine current state of the vehicle
relative to the ground. In this paper, “Multi Axis Load
Sensor-Integrated Hub Bearing” NTN has developed a
new bearing to address this challenge and it isdescribed hereunder.

2.1 Development concept
There are two approaches for ensuring safety
during operation of automobiles: active safety, wherea symptom of an abnormal situation is detected andthe possible hazard is avoided, and passive safety,
that helps protect human bodies in case of collisionand other accidents. Typical means for active safetyinclude ABS (anti-lock braking system) and ESC (antiskidding system); and those for passive safety include
seat-belts and air-bags. Focusing on improvements inactive safety functions, NTN determined the concept
for this development project, which is, detectingchanges in frictional force on tires based on
information from sensors situated nearest to the roadsurface (the sensors are integrated in the hub
bearings). Vehicle information that is utilized for active
safety is summarized in Table 1.

2.2 Structure
Fig. 1 shows the appearance of the new hub
bearing NTN has developed for the above-mentioned
purpose. The detection principle of this new bearing
product is schematically illustrated in Fig. 2. A strain
sensor is installed to the outer ring (fixed ring) of a hub
bearing: when the car is travelling, a strain occurs on
the outer ring, and the strain sensor detects the strain,
thereby the active safety system can detectlongitudinal force (Fx), lateral force (Fy) and vertical* Automotive Business HQ. Axle Unit Engineering Dept.
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Hub Bearing with Integrated Multi-axis Load Sensor

force (Fz) working between the road surface and tire;
independent of the motion of the suspension.
When a load is applied to the hub bearing,
infinitesimal deformation occurs on the outer ring of
the hub bearing, wherein the pattern of thedeformation can vary depending on the direction ofloading. This infinitesimal deformation is detected byfour sensors (situated in upper, lower, right and left
locations), and then the resultant sensor signals are
subjected to arithmetic operations in order todetermine the magnitude of the loads in three
directions. Also, since the load-induced deformationon outer ring is detected, the load can be detectedbased on the weight of vehicle even when the tires areat a standstill (locked).

2.3 Bearing types for load sensor-integrated
hub bearing
Fig. 3 illustrates the types of load sensor-integrated
hub bearings. Installed to the fixed outer ring of GEN2and GEN3 flanged hub bearings whose inner ring canfreely rotate, the load sensor detects the load acting
on the bearing.

Table 1 Vehicle information applicable to active safety system
Characteristic to be detected

SensorMounting location

Response for change detection

Vehicle velocity detection

ABS (wheel speed) sensor

Hub bearing (axle)

Real time

Steering angle detection

Angle sensor

Steering system

Real time

Angular velocity detection

Yaw rate sensor

Onboard (sprung)

Minor delay

Acceleration detection

G sensor

Onboard (sprung)

Minor delay

Tire load (change in μ) detection

Multi axis load sensor

Hub bearing (unsprung)Real time

Infinitesimal deformationon outer ringFz

Database

Fy
Arithmetic operation
for load

Fx
Fx = Longitudinal force
Fy = Lateral force
Fz = Vertical force

Detection of forces on
three axes (Fx, Fy, Fz)
Load detection algorithm

Fig. 2 Load detection scheme

Fig. 1 Load sensor-integrated hub bearing

GEN2

GEN3

Drive ring

Drive ring
Flanged outer ring

Flanged outer ringSensor mount

Flanged outer ringSensor mount

Knuckle side

Wheel side

Structure

Sensor mount

Driven ring

Inner ring

Fig. 3 Bearing types for load sensor-integrated hub bearing-59-

NTN TECHNICAL REVIEW No.79（2011）

point is nearly impossible when the detection point isaxially deviated from a location where the rollingsurface is situated directly beneath it or the detection
point is situated on the non-load side on outercircumference and if this variation in strain derives
from this location and passage of rolling elements.
Additionally, the two types of strain in Fig. 5 eachexhibit unique characteristics depending on themagnitude and direction of loading.
Therefore, a plurality of strain sensors capable ofdetecting these two types of strain are installed on theouter circumference of the outer ring in order to allow
the load on outer ring to be accurately estimated.

3. Load Detection Algorithm
3.1 Strain detection system
When a load acts on the outer ring, the magnitudeof strain at a given location on the outer ring varies
depending on the magnitude of external load, and at
the same time, is affected by passing rolling elements.
The magnitude of strain on a running bearing that isbeing affected by external load varies when rolling
elements pass by, and is maximized at the very
moment where a rolling element passes by a given
strain detection point and is minimized at a moment
where the same strain detection point is situated at
midpoint between the two neighboring rolling elements(Fig. 4). In other words, even if the magnitude ofexternal load remains unchanged, the magnitude ofstrain will vary depending on the positional
relationship between a strain detection point and therolling elements.
As can be understood from the discussion above,
strain occurring at a given detection point under a loadof certain magnitude essentially consists of a constantcomponent from external load and a varying
component from passing rolling elements (Fig. 5).
Incidentally, strain at a given detection point on theouter circumference of the outer ring, derived from anexternal load, can be detected regardless of the axial
location of that detection point. However, variation in
the magnitude of strain at a particular strain detection

External load

Maximumstrain3.2 Shape of sensor and signal processing system
To be able to amplify infinitesimal deformation on anouter ring, the strain sensor used is a plate typesufficiently thin relative to the outer ring. The strain
detection area of the sensor includes necks (Fig. 6)
while the mount for the sensor has slots so that
deformation around these slots concentrates on thestrain detection area to improve detection sensitivity ofthe sensor.
As previously described in Sec. 3.1, the strain
sensor also detects strain from the rolling motion of
the rolling elements; consequently, even when aconstant load is applied to the outer ring, the level ofsensor output signal can vary depending on thelocation of rolling elements. For example, when the
bearing is running, variation in strain within one cycle
owing to passing rolling elements can be determinedby arithmetic operation to calculate the magnitude ofstrain (strain from external force) corresponding to themagnitude of external force. In contrast, when the
bearing is at a standstill, the strain component of theexternal force cannot be readily separated from themagnitude of strain detected by one strain gage alone.
To address this issue, NTN has adopted a uniquesensor structure characterized in that one strain
sensor has two detection areas as shown in Figs. 6
and 7. These detection areas are situated with aspace 1/2 as large as the pitch between adjacentrolling elements (balls), and two outputs from thesedetection areas are summed up in order tocompensate for variation in the output that is governedby the location of rolling elements. Consequently, it is
now possible without adopting an outside sensor or
the like to estimate the load on the outer ring evenwhen the tire is at a standstill.

External load

Strain detection point

Rolling element is directly
beneath detection point.

MinimumstrainDetection point is at a midpoint
between rolling elements.

Strain owing to rolling
motion of rolling element

Fig. 4 Influence of location of rolling elements

Strain onouter ring

Strain owing toexternal force

3.3 Load estimation method
First, sensor output and input loads under expecting
driving conditions for both simulator testing and roadtesting are measured; thereby, a database for the load

Fig. 5 Strain occurring on outer ring-60-

Hub Bearing with Integrated Multi-axis Load Sensor

estimation matrix is developed. An arithmetic
operation for matrix input conditions includes various
scenarios and combinations for various driving
conditions estimated from sensor output rather thansimple sensor outputs themselves. Thus, adoption of
NTN’s unique calculation technique helps offer a load
estimation approach that boasts higher accuracy
compared with conventional techniques.

4. Result Assessment by Road Test
4.1 Test conditions
NTN's multi axis load sensor-integrated hub
bearings were incorporated in the four wheels on thetest vehicle that was subjected to road testing; thereby
load detection accuracy and response of thesesensors were assessed. For assessing load detection
accuracy, output of a six-component force transducer
installed in the tire was taken as “tire-to-road surface”
load, and was compared with the output of a multi-axisload sensor. For assessing sensor response, output of
the onboard yaw rate sensor was compared with
output of the multi-axis load sensor. Table 2
summarizes the specification of the sensors mounted
on the test vehicle.

Strain gage

Table 2 Load sensor specification

Neck

Characteristic

1/2 ball pitch interval

ValueLongitudinal force:Fx

Fig. 6 Strain sensor

Lateral force:Fy

Detection load

Vertical force:Fz
Detectable load range

Sum of outputs from two strain gages

Sensor output V

Targeted estimation
accuracy
Sampling frequency
Response

Fx

±10kN

Fy

-4kN〜10kNFz

0〜10kN

Fx±5%FS(±1kN)

Fy±5%FS (±0.7kN)

Fz±5%FS(±0.5kN)
300Hz
30 ms max.

Subject of comparison for accuracy Six-component force transducer* FS: Full Scale
Time sStrain gage

4.2 Assessment of the load detection accuracy

Strain gage

Road testing was performed using a slalom mode,lane change mode and acceleration/deceleration
mode; thereby the load detection accuracy of thesensor in each road mode was determined. As
representative examples, road test results of theslalom mode and lane change mode are graphicallyplotted in Figs. 8 and 9. These graphs show variationin the outputs from the six-component force
transducer and multi-axis load sensor. At each time
point, output error of multi-axis load sensor relative tothe output of the six-component force transducer wasdetermined.
In the slalom test whose result is graphically plottedin Fig. 8, the error in component forces in three
directions (Fx, Fy, Fz) each fall in the targeted range(±5%FS) relative to the outputs from the six-component
force transducer.
In quick acceleration and lane change testing,results in Fig. 9, the change in three component

Outer ring
Directionof rolling
Rolling element
Ball pitch
1/2 ball pitch

Fig. 7 Signal processing scheme

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10
8
6
4
2
0
-2
-4
-6
-8
-10

10

Load sensor
Six-component force
transducer

0

1.5

3

Fx load kN

Fx load kN

NTN TECHNICAL REVIEW No.79（2011）

3.5

4

8
6
4
2

Load sensor
Six-component force
transducer

0
-2
-4
-6
-8
-10

4.5

Time sQuickacceleration
Lane change
0

3

6

9

Load sensor

6

Six-component force
transducer

4
2
0
-2
0

1.5

3

3.5

4

4.5

Time s8

Load sensor

6

Six-component force
transducer

21

15

18

21

15

18

21

4
2
0

-4

0

3

6

9

Load sensor

812Time sSix-component force
transducer

6

10

4

8

Load sensor

Fz load kN

Fz load kN

18

-2

10

2
0

15

10

Fy load kN

Fy load kN

8

-4

12

Time s

10

0

1.5

3

3.5

4

4.5

Time sSix-component force
transducer

6
4
2

Fig. 8 Result of slalom test

0
0

3

6

9

12

Time sFig. 9 Result of quick acceleration/deceleration and lanechange tests

20

2

Yaw rate sensor

10

Yaw rate deg/sec

4.3 Assessment of response
In Sec. 4.2, the load detection accuracy andresponse capability for variation in load was verified
for the six-component force transducer built into thewheel. In this section, output from the onboard yaw
rate sensor is compared with output from the multiaxis load sensor in the hub bearing in order to verify
response of the load sensor. The graphical plotting in
Fig. 10 illustrates changes in outputs of the yaw rate
sensor and the multi-axis load sensor that occur when
the steering wheel on a car running straight is turned
quickly. It should be understood that compared withthe response of the yaw rate sensor, the response of
the multi-axis load sensor is approximately 10 msquicker. This difference appears to coincide with atime span from the occurrence of lateral force on thetire to detection of this force by the onboard yaw rate

1

0

0

-10

-1

-20

-2

-30

-3

100ms
-40

-4

-50
-60
11.4

-5

Load sensor
11.5

11.6

11.7

11.8

11.9

12.0

Time sFig. 10 Comparison of responses

-62-

-6
12.1

Load sensor Fy N

forces and output from the multi-axis load sensor issimilar to that of the output from the six-component
force transducer: it has been proven that the multi-axisload sensor positively detects a quick change in loadthat can result from quick acceleration/deceleration orviolent lane change.

Hub Bearing with Integrated Multi-axis Load Sensor

Reference

sensor included in the sprung mass. In other words, if
the car is traveling at 100 km/h, this time difference
coincides with a thinking distance of approximately 3m.
On the basis of these findings, NTN believes thatthe ability to estimate tire-to-road status with a multiaxis load sensor will help ensure driver safety duringsituations including hazard avoidance, and that thisarrangement will provide useful vehicle information
while improving ride comfort by controlling the
behavior of the vehicle before the driver feels
uncomfortable motion of the vehicle.

1) Norimatsu, T., Nishikawa, K., "Development of Multi
Axis Load Sensor Integrated Hub Bearing," Journal of
the Society of Automotive Engineers of Japan, Vol. 64,
pp. 104-105, 2010.
2) Norimatsu, T., "Technical trends in hub bearings withintegrated sensors," The Tribology, No. 280, Shinjusha
Co., Ltd., pp. 42-43, 2010.

5. Conclusion
To help assist in further sophistication of the vehicle
control technique, NTN has developed a “multi-axisload sensor-integrated hub bearing” that can detectvariation in frictional force on the tire in a real timebases at a location the near road surface.
In the presented road test, it has been verified that
the NTN’s multi-axis load sensor provides output
equivalent to that obtained from a six-component force
transducer, and boasts better response comparedwith onboard sensors included in the sprung mass
such as a yaw rate sensor. NTN will seek thepossibility of vehicle control using an intelligent corner
module that comprises a load sensor alone or a load
sensor in conjunction with a “high resolution sensor2)”
or the like, in an attempt to adopt improved vehicle
control systems to further enhance vehicle stability.

Photo of author

Kentaro NISHIKAWA
Automotive Business HQ
Axle Unit Engineering Dept.

-63-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Hub Bearing for Severe Environments

Yasushi SHIBATA*

With the increase of automobile demand in emerging markets, thenumber of cars driven in severe environments, such as on unpaved roads
or in cold and snowy weather, is increasing.
A hub bearing for use in such severe conditions is introduced in thisreport.

Focusing on a design concept for hub bearing
design that can withstand these severe environments
and promised designed bearing life, NTN has
executed the following development projects:¡Development of an improved seal for muddy water
resistance¡Development of bearing internal structure for
improved load capacity (impact resistance)¡Development of grease that helps improve fretting
resistance of bearing components
A typical example of the structure of a hub bearing---topic of this report --- is shown in Fig. 1.
Note that the new development (severe environmentcapable hub bearing) is equivalent with NTN'sconventional product in terms of bearing life and torque.

1. Introduction
Recently, demands for cars have been everincreasing in many emerging countries and many ofthese cars are traveling on poor roads including
unpaved roads. In this context, automotive hub
bearings are expected to meet increased muddy water
resistance and load bearing capacity.
During long distance transportation of automobiles
on railway freight cars in cold weather, raceway
surfaces on automotive hub bearings can develop
fretting owing to shocks that occur when freight cars
pass rail joints: as a result, the hub bearings candevelop abnormal noise and premature failure.
This paper hereunder describes NTN's unique "hubbearing for severe environment (or, severe
environment-capable hub bearing)" that helps solve
the above-mentioned problem and contributes toreliability of automobiles.

Outer side

Inner side
Outer seal

Hub bolt

2. Development Concept and Structure
Typical severe environments which automotive hub
bearings can experience include:¡Ingress of muddy water into hub bearing as cars
more frequently travel on muddy road.
¡ Excessive loading onto the hub bearings as cars
travel on poor roads and/or carry excessively heavy
cargos.
¡ Possible fretting wear resulting from long distancerailway transportation on freight cars in cold
weather.

Inner seal

Inner ring
Hub ring

Rolling element
Outer ring

Fig. 1 Typical structure of hub bearing* Automotive Business HQ. Axle Unit Engineering Dept.
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Hub Bearing for Severe Environments

optimized contact width and bearing pressure
between the seal lip and hub ring. Similar
improvements were made to the inner seal as well forblocking muddy water that possibly runs along theouter ring and penetrates the seal.

3. Development Specification
3.1 Improved muddy water resistance
As illustrated in Fig. 1, there is an unavoidable
difference in structure between the inner seal of a hub
bearing and outer seal. Therefore, to be able toimprove muddy water resistance of a hub bearing, thestructures of the inner seal and outer seal have eachbeen optimized.

3.2 Improved load capacity
When automotive tires roll on unpaved road or ride
on a curb, an abnormally large load will momentarily
act on the hub bearings. When greater load acts on a
hub bearing, the contact area between the raceway
surface and rolling element shifts toward the shoulderedge of the raceway as shown in Fig. 6; thereby
causing the rolling elements to ride on the raceway
shoulder edge. Consequently, stress concentration
can occur on the shoulder edge of the raceway andimpact dents tend to occur in this area.
Consequently, when rolling elements roll over the
so-formed fretting marks, abnormal noise andvibration can occur as schematically illustrated in Fig.
7.
To inhibit the occurrence of impact denting, astraight section has been added as shown in Fig. 8
toward the tangential direction from a conventional
raceway shoulder so that the height of raceway
shoulder is increased. With this arrangement, thetendency of the rolling elements riding onto raceway
shoulder is limited to help reduce stress concentration
at raceway shoulder.

(1) Inner seal
As shown in Fig. 2, the structure of a conventional
inner seal consists of a side lip and radial lip A eachintended to block ingress of muddy water fromoutside, and a radial lip B for containing grease
prefilled in the hub bearing.
To improve muddy water resistance, the newly
developed inner seal shown in Fig. 3 has side lip B inplace of conventional radial lip A, and at the sametime, contact width and bearing pressure between theseal lip and slinger have been optimized.

(2) Outer seal
As shown in Fig. 4, the structure of a conventional
outer seal consists of a side lip and intermediate lipeach intended to prevent ingress of muddy water fromoutside, and a radial lip for containing the prefilledgrease.
To improve muddy water resistance, NTN’s newly
developed outer seal shown in Fig. 5 featuring

Side lipHub
ring

Slinger
Side lip
Radial lip A

Outer
ring

Intermediate lip
Rolling
element

Radial lip
Radial lip B

Fig. 2 Structure of a conventional inner seal

Fig. 4 Structure of a conventional outer seal

Weir
Side lip A
Side lip B

Hub
ring

Outer
ring

Slinger
Rolling
element

Radial lip

Fig. 3 Structure of the newly developed inner seal
Fig. 5 Structure of the newly developed outer seal

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NTN TECHNICAL REVIEW No.79（2011）

3.3 Fretting resistance enhancing grease

Normalstate

Raceway
surface
Raceway
shoulderDuring long distance transportation of automobiles
on railway freight cars, minor oscillations will occur
between the raceway surfaces and the rolling
elements in automotive hub bearings; thereby fretting
wear can result as schematically illustrated in Fig. 9.
In minor oscillation mode, a portion of grease
between the raceway surface and rolling elements will
be pushed out of the contact path; consequently,
metal-to-metal contact will occur, possibly leading tofretting wear. In a cold weather situation, the fluidity ofgrease will be further deteriorated; thereby fretting
wear can occur more frequently.
To address this problem, NTN has adopted improved
grease additives to help generate a reactionary film
between the raceway surface and rolling elements inaddition to an oil film in order to inhibit the occurrence of
metal-to-metal contact situation.

High loadingstateState of contact between
rolling elements andraceway surface

State of contact between
rolling elements andraceway surface

Stressdistribution

Stressdistribution

Oscillation on automobile that occurs
when freight car passes rail joints.

Peak location

Fig. 6 State of raceway shoulder
Automobile
Rotation
Tire
Revolution

Minor oscillation occurs
between raceway surface
and rolling elements.

ImpactFretting wear
Shoulder edgeon groove

Occurrence of noiseand vibration

Fig. 7 Mechanism of occurrence of noise and vibration
Bottom on groove

Fig. 9 Mechanism of occurrence of fretting wear
Outer ring

4. Assessment Test
Shape of a conventional raceway
surface (arc shape)

Increasedshoulder height

To verify the advantages of NTN's "severe
environment capable hub bearing", in terms of muddy
water resistance, load capacity and fretting resistance,
this bearing type has been subjected to rig
assessment tests (1) through (3) summarized below:(1) Muddy water resistance: Muddy water resistancetest for seal
(2) Load capacity (impact resistance): Static high-loadbearing test(3) Fretting resistance: Fretting wear test with newly
developed grease

Tangential straight section
starting at shoulder ofconventional raceway surface

Fig. 8 New raceway design-66-Hub Bearing for Severe Environments

4.2 Improved load capacity

To assess the muddy water resistance of NTN's
severe environment-capable hub bearing, the bearing
samples were repeatedly run and stopped with theirseals submerged in muddy water up to their axial
center.
The degree of mud water resistance for each hubbearing sample has been determined based on thenumber of cycles run at which muddy water has
overcome the seal.
As shown in Fig. 10, it has been verified that boththe newly developed inner seal and outer seal feature
muddy water resistance twice as great compared withconventional seals.

To assess load capacity (impact resistance) of
NTN’s newly developed hub bearing, bearing samples
were subjected to a static load equivalent to acornering force of 2G or greater on a vehicle tosimulate lateral collision to curb. Thereby, stress
distribution was analyzed and impact testing was
performed for the newly developed hub bearing.
As illustrated in Fig. 11, the analysis result showsthat the newly developed shape of the raceway
surface positively reduces stress occurring on the
raceway shoulder.
In addition, as a result of the impact test, it has beenlearned that the depth of impact dents on the raceway
shoulder using the improved shape is approximately
1/4 as deep compared with the conventional raceway
shape as shown in Fig. 12.

4[Inner seal]
3

2

Conventional
shape

1

0

Newly
developed
designStressgenerated

Conventional
designRatio of muddy water resistance(ratio in terms of number of cycles)

Newly developed
shape

Higher4[Outer seal]
Lower

3

Raceway
shoulder2

Bottom ofgrooveRaceway
shoulderBottom ofgrooveFig. 11 Result of stress distribution analysis on raceway
surface10Conventional
designDepth of fretting on raceway shoulder(μm)

Ratio of muddy water resistance(ratio in terms of number of cycles)

4.1 Muddy water resistanceNewlydeveloped
design※Conventional design is taken as 1.

Fig. 10 Muddy water test result

20
18
16
14
12
10
8
6
4
2
0

Conventional
shape ofraceway surface

Newly developed
shape ofraceway surface

Fig. 12 Typical result of fretting test

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NTN TECHNICAL REVIEW No.79（2011）

4.3 Fretting resistance

5. Conclusion

Ratio in terms of fretting wear depth

A fretting wear test was performed by alternately
applying lateral load and longitudinal load to the hub
bearing samples, simulating a vibration patternoccurring in automobiles transported by railway.
As shown in Fig. 13, depth of fretting wear on
raceway surface lubricated with conventional grease
under unusually low temperature is as great as 10times or more compared to when the bearing is at
operating temperature. In contrast, since the additivein the newly developed grease helps generate areactionary film between the steel balls and raceway
surface, severity of fretting on the raceway surface
lubricated with the improved grease under lower
temperature is equivalent to raceway surface
lubricated with the same grease under operating
temperatures.
Additionally, NTN has executed test requirements
the bearing needs to meet, other than fretting
resistance, which include bearing life, high speed
grease life, and running torque, thereby NTN has
found that the performance of NTN’s newly developed
grease is equivalent to or better than conventional
grease.

NTN has developed a unique “severe environmentcapable hub bearing” that boasts improved muddy
water resistance, improved load capacity, andimproved cold weather fretting resistance. These
improvements allow for increased performance insevere environments were the road conditions arepoor and vehicles are transported by rail overextended distances. As the demand for automobiles
continues to increase in emerging countries,
automobiles are expected to be used in increasingly
demanding environments. To address this situation,
NTN will further improve its technology described inthis report.

Reference
1) Shibata, K., Norimatsu, T., NTN Technical Review No.
75, p. 31, 2007.
2) Nagatani, H. and Imou, A., "Contact Pressure andShear Stress Analysis on Conforming Contact
Problem," Journal of Advanced Mechanical Design,
Systems, and Manufacturing, Vol. 2, No. 6, pp. 10551066, 2008.

20

15

Conventional grease
10

5

Newly developed grease10

Ordinary
temperature

Lower
temperature※Value with conventional grease at ordinary temperature is taken as 1.

Fig. 13 Result of fretting wear test

Photo of author

Yasushi SHIBATA
Automotive Business HQ
Axle Unit Engineering Dept.

-68-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Lightweight and High-efficiency Drive Shafts
for Rear-wheel-drive Cars

Tatsuro SUGIYAMA*
Yuichi ASANO*

A lightweight and high-efficiency drive shaft designed
exclusively for rear-wheel-drive cars with a mass reduction by
16% and a torque loss rate reduced by 40% in comparison
with conventional products.

Recently, needs for more compact, lighter, more
efficient drive components have been mounting to
reduce fuel consumption in addressing environmental
conservation. This trend is also true with luxury cars oflarger engine displacement.
Generally, RWD cars use CVJs whose design
specification is identical to that of CVJs for FWD cars.
Nevertheless, NTN has newly developed highperformance drive shaft dedicated to RWD cars. Thisdrive shaft is intended for luxury RWD cars that require
an amply sized (high load capacity) drive shaft: thisproduct is compact, light-weight, much more efficient
rear wheel-dedicated drive shaft that contributes tolighter car design and improved fuel economy.
This special rear wheel drive shaft consists of two
sets of newly developed compact, light-weight, highly
efficient cross groove type plunging CVJs that arecoupled with each other via a light-weight hollow shaft.
The structure and features of this new product are
hereunder described.

1. Introduction
Constant velocity joints (hereinafter referred to as
CVJ) are often used for automotive drive shafts. Frontwheel-drive (FWD) cars such as a one shown in Fig. 1
use fixed CVJ (for example, NTN’s EBJ) on the tireside because of greater operating angle. In the engine
side, the CVJ must be capable of sliding in axial
direction to compensate for suspension movement(for example, NTN’s ETJ) is used.
On the other hand, rear-wheel-drive (RWD) cars
can use a tire-side CVJ whose operating angle is not
greater than 23˚. Thus, RWD cars can adopt mounting
procedure for fixed and plunging CVJs reverse to thatof FWD cars, or may utilize drive shafts whose bothsides are capable of plunging.

Tire

2. Structure of Newly Developed
Driveshaft

Wheel
bearing

Fixed CVJ
(EBJ)

Shaft

The structure of NTN's newly developed drive shaft
is shown in Fig. 2
Both ends of this drive shaft are each equipped witha cross groove type plunging CVJ featuring improved
efficiency and smaller size through increasing thenumber of balls as compared with conventional crossgroove CVJ’s, wherein both CVJ’s are linked with alight-weight hollow shaft not undergoing any drawing
processes.

Plunging CVJ(ETJ)Fig. 1 Construction of drive shaft for FWD cars

* Automotive Business HQ. C.V. Joint Engineering Dept.
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NTN TECHNICAL REVIEW No.79（2011）

Shaft
Light-weight hollow type
Plunging CVJ (tire side)
ELJ stem type

Plunging CVJ (engine side)
ELJ disk typeFig. 2 Newly developed construction of drive shaft for RWD cars

Various structures are available for installing CVJ to
car bodies. On an example in Fig. 2, the mountsystem in tire side of CVJ is stem mount system
where the splined shaft end of CVJ is fitted into thecounterpart (hub) and the mounting system in the
differential side of CVJ is a disk mount system where
the shaft end of differential side is bolted down to thecompanion flange; thereby, stem specification may be
applied to both drive shaft ends.
The two CVJs on both ends of drive shaft may be
identical in terms of their internal structure; thereby,
the plunging CVJs at both ends of drive shaft can bothaccommodate axial displacement.
Compared with conventional cross groove CVJ(NTN's LJ), the NTN's newly developed CVJ has anincreased number of balls of smaller diameter. This
new CVJ design has contributed to higher efficiency
and much compact size of drive shaft.
This unique CVJ has been designated "ELJ" that
constitutes a part of NTN's E series constant velocity
joint products boasting light weight, compact size andhigher efficiency.
In addition, the light-weight hollow shaft for coupling
the two ELJ CVJs is joined to the inner rings by amply
sized splines; thereby, the thin-walled shaft having
near-uniform bore diameter and outside diameter overthe entire shaft length can be designed, and thisfeature contributes to realization of light-weight drive
shaft. This drive shaft is very unique in that it does not
require drawing at the shaft ends that is necessary forstandard splined joint on a conventional hollow shaft(refer to Figs. 2 and 8).

3. High-Efficiency Plunging CVJ (ELJ)
3.1 Construction and features of newly
developed CVJ
NTN's conventional cross groove plunging CVJ "LJ"
has six cross grooves for rolling balls on the inner ringand outer ring; wherein, these grooves are oriented
inclined toward the circumferential direction and two
adjacent grooves are inclined in reverse directions(refer to Figs. 4 and 5). The balls are held at theintersection points between cross grooves on innerring and outer grooves (grooves on the inner ring are
inclined in the direction opposite to that of grooves onouter ring), all the balls are situated on a common
plane in a bisector with the cage; thereby, the CVJ
can run at constant speed while the inner ring canslide in the axial direction.
Fig. 3 shows an exploded view of NTN's newly
developed cross groove CVJ "ELJ", and Fig. 4 gives
an exploded view of conventional cross groove CVJ"LJ". Compared with "LJ", "ELJ" features increased
number of balls of smaller diameter, and inclination ofcross grooves on inner ring and outer ring is smaller;
thereby the ELJ CVJ is capable of maximum operating
angle equivalent to that of conventional CVJ "LJ" whileachieving compact size, lighter weight and more
efficient torque transmission.
Table 1 summarizes characteristics of NTN's newly
developed cross groove plunging CVJ "ELJ" ascompared with those of conventional cross groove
CVJ "LJ". Fig. 5 schematically illustrates difference
between the newly developed cross groove CVJ andconventional cross groove CVJ in terms of inclinationof their cross grooves by referring to cutaways of theirdisk type outer rings.

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Lightweight and High-efficiency Drive Shafts for Rear-wheel-drive Cars

3.2 Reduced size and weight with "ELJ"As previously mentioned, "ELJ" has achieved
smaller size through size reduction and increase in
number of balls and decrease in inclination of crossgrooves for rolling balls. At the same time, smaller ballsize allows the wall thickness of the inner ring to begreater; consequently, the diameter of splined bore of
inner ring to be joined to the splined shaft end can begreater. As a result, the axial dimension of inner ringcan be further reduced; thus the weight of inner ringhas been much reduced.
In Fig. 6, a cutaway of the newly developed stem
outer ring type ELJ109 (109 is designation of NTN'sdrive shaft) is compared with that of conventional
design. Compared with conventional design, theoutside diameter of ELJ 109 is approximately 9%
smaller and weight reduction with the joint alone has
amounted to approximately 22% (2.63 kg to 2.05 kg).

Outer rig
Ball
Cage

Inner ring

Shaft

Fig. 3 Newly developed cross groove type CVJ "ELJ"

Outer rig
Ball

SplinesCage Inner ring

Shaft
φ89

Fig. 4 Conventional cross groove type CVJ "LJ"
Newly
developed

Table 1 Comparison of characteristics of ELJ with those of LJ
Newly developed ELJ Conventional LJ

Max. operating angle

10

6

Low

High

23 deg23 degConventional
φ98

Number of balls
Inclination ofcross groove

Fig. 6 Comparison of cutaway of ELJ with that ofconventional cross groove type CVJ

3.3 Efficiency enhancement for ELJ
Smaller ball size, increased number of balls, andreduction in inclination of cross grooves on inner ringand outer ring each help improve efficiency of CVJ.
Reduction in inclination of cross grooves helps reduce
axial sliding friction on the balls, and decrease the shiftof balls; consequently, torque loss is much decreased.
Fig. 7 shows the result of torque loss reduction ratiowith ELJ with the operating angle of CVJ kept at 5degrees. Compared with conventional "LJ", the torqueloss with the newly developed "ELJ" is approximately
40% smaller.

Cross groove forrolling ball
Newly
developed

Lower inclination

Higher inclination

Conventional
Torque loss ratio

Reduction of approx.
40%

Fig. 5 Schematic illustration of inclination of cross groove
for rolling balls

Conventional LJ

Newly developed ELJ

Fig. 7 Torque loss reduction ratio with ELJ-71-

NTN TECHNICAL REVIEW No.79（2011）

hollow shaft achieved weight reduction of approximately
18%. Thanks to this weight reduction, hollow drive
shaft may have greater wall thickness to enhancerigidity of the hollow shaft provided that increase inwall thickness does not lead to weight of drive shaft
greater than that of conventional design.

4. Light-weight hollow shaft
Conventional hollow shafts used in drive shafts have
limitation in size of splined section because as shownby the lower example in Fig. 8, a smaller diameter isadopted for shaft end while maintaining sufficient wall
thickness for the inner ring. Since necessary strength
needs to be maintained for shaft end even with asmaller diameter, the shaft end is subjected to drawing
to provide greater wall thickness for this area.
The newly developed drive shaft is dedicated to
RWD applications, and will not be operated at greater
operating angle. Furthermore, adoption of the newly
developed “ELJ” helps enlarge the splined section forjoining to the inner ring as previously explained in Sec.
3.2; consequently, as shown by the upper example inFig. 8, the newly developed hollow shaft does notneed drawing on either end. This is because the
NTN’s newly developed drive shaft technology allows
the diameter of splined section on shaft end to belarger; consequently, sufficient mechanical strength
can be provided for both ends of the hollow shaft
without modification of both ends to increase wall
thickness.
NTN's new hollow drive shaft features thinner wall
thickness, with its bore diameter and outside diameterremaining uniform over nearly entire longitudinal
dimension.This feature helps to further reduce theweight of drive shaft. For comparison purpose, ahollow shaft whose outside diameter at its midpoint isequal too the newly developed drive shaft, with bothends subjected to a drawing process such that its bothends have strength equivalent to the newly developed
drive shaft; then the weight of this comparison hollow
shaft was compared with the new drive shaft.
Consequently, it has been found that the NTN’s new

5. Conclusion(1) NTN's newly developed rear wheel drive shaft
achieves approximately 16% of weight reduction*1
by adoption of "ELJ" + thin-walled hollow shaft +"ELJ".
(2) Adoption of "ELJ" realizes approximately 40%
reduction in torque loss*2 (higher efficiency)(3) Adoption of "ELJ" realizes approximately 9% of
reduction in outside diameter*2 (outside diameterof stem outer ring type).
(4) Adoption of thin-walled hollow shaft that has notundergone the drawing process on either end
leads to further weight reduction.
*1 Comparison in weight of drive shaft assembly withconventional design LJ + hollow shaft (drawn shaft ends).
Total of approximately 16% weight reduction by stem outerring type CVJ (Sec. 3.2, approximately 22% weight reduction)+ hollow shaft (Sec. 4, 18% weight reduction) + disk outer ringtype CVJ.
*2 Comparison with conventional LJ.

Compared with conventional LJ, NTN’s newly
developed drive shaft (ELJ) dedicated to rear wheel
drive cars has achieved size reduction, lighter weightand higher transmission efficiency. As people aregoing to be more eco-conscious, NTN will apply its
improved drive shaft technology to automobiles thatneed higher efficiency in transmitting driving power.

Light-weight hollow
shaft, not drawn(newly developed)
Not drawn

Photo of authors
Light-weight hollow
shaft, drawn(conventional)

Drawn
Drawn

Fig. 8 Comparison of hollow shafts in terms of shape-72-

Tatsuro SUGIYAMA

Yuichi ASANO

Automotive Business HQ
C.V. Joint Engineering Dept.

Automotive Business HQ
C.V. Joint Engineering Dept.

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Ball Screw Unit for Automotive Electro-actuation

Koji TATEISHI*

In the automotive market, numerous new hybrid cars and engineswith low fuel consumption and low emissions have been developed toreduce CO2 emissions and to make their exhaust cleaner. At the sametime, many projects are being undertaken to achieve greater car
amenities and safer driving by applying quick and reliable electricmotor drives utilizing ball screws.
NTN started to develop new ball screw units and actuators in the
1990s. Now there are some ball screw products that are modularizedwith peripheral parts. This article introduces the results ofdevelopment and the features of these ball screw and actuator units.

1. Introduction

2. Structure and Features of Automotive
Electromechanical Ball Screw

Fuel economy of automobiles was dramatically
improved when hydraulic pumps for auxiliary engine
systems were superseded by electromechanicalactuation system (adoption of driving system that are
based on electric motors) that run only when theauxiliary engine systems need to operate. Control-bywire systems for controlling automobiles, such as shiftby-wire or throttle-by-wire, have recently become morecommonly commercialized. In the present day
European car markets, brake-by-wire control
technology is being commercialized.
NTN has developed and is currently marketing
unique quick-response, high-thrust automotive
electromechanical ball screw units that can be used inelectromechanical control-by-wire actuators for theabove-mentioned an automotive applications.
This paper hereunder describes automotive
electromechanical ball screw unit andelectromechanical ball screw actuator (hereinafterreferred to as "actuator unit" ) NTN has developed.

Being common with any rolling bearings, ball screw
units include rolling mechanism involving balls. Thus,ball screw units have been often used on industrial
machines as machine elements that can efficiently
convert rotary motion into linear motion.
However, this useful mechanical component has
failed to achieve wider support in the automotive
component market because the conventional ballscrew unit design does not meet productivity needed
for mass production and at the same time the ball
circulator---a key mechanism in ball screw--- may fail tosatisfy durability requirement of 10 operating years or
0.30 million km of distance traveled. So that ball screw
units can be applied to automotive applications, NTN
has begun development activities that are to be
described in the following sections.

2.1 Improved reliability for ball circulator
Previously, a return tube type ball circulator has
been widely used because of the relative ease ofdesign and manufacture. Fig. 1 illustrates a typicalexample of a return tube type ball circulator.
Because this type of ball circulator uses a ballpicker it needs improved mechanical strength, NTN
adopted a deflector type ball circulator for automotive* Automotive Business HQ. Axle Unit Engineering Dept.
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NTN TECHNICAL REVIEW No.79（2011）

applications. Fig. 2 illustrates a typical example ofdeflector type ball circulator.
NTN's deflector is first inserted into the bore side of
the nut and then fixed to the nut by a staking technique,thereby the deflector boasts improved reliability andcompact size.

Conventional induction
hardened productBall picker
Return tube
Newly developed induction
hardened productFig. 3 Appearance of thread groove on screw shaft

Ball

Fig. 1 Return tube type ball circulator

Deflector

Wing

Fig. 4 Appearance of thread groove on nut

Movement of ballof the nut was a challenge in particular when seeking
high precision finish. By adoption of a CNC-controlled
high-precision cutting technique, NTN has
successfully formed thread grooves on the bore
surface of the nut without involving grinding operation,
wherein the so-formed thread groove boasts sufficientdurability and operability. Fig. 4 shows theappearance of thread grooves on the bore surface ofa nut achieved by NTN’s unique high-precision cutting
process.

Bore of nut

Fig. 2 Deflector type ball circulator

2.2 Improved productivity
Previously, the threaded groove of the ball screw
had been formed by grinding with a grinding wheel. To
be able to reduce consumption of grinding wheels anddecrease machining energy used through size
reduction of machining equipment, NTN has
developed a ball screw machining technique that does
not involve a grinding operation.
To generate the thread groove on a screw shaft, aprecision thread rolling process that uses roller dies is
adopted to form thread grooves in one pass. Then, thescrew shaft is subjected to a newly adopted induction
hardening process that does not generate scale,thereby machine cycle time has been reduced. Fig. 3
shows the appearance of the thread grooves on thescrew shafts: one example having undergone theconventional induction hardening process and theother example having undergone NTN’s newly
developed induction hardening process.
Generation of the thread groove on the bore surface

3. Application of Electromechanical Ball
Screw Unit to Electromechanical Actuator
NTN's experience in the development work forelectromechanical ball screw units and actuator units
that NTN has already marketed is hereunder
described.

3.1 Ball screw unit for automotive automatedmanual transmission
Automated manual transmission (AMT), which
automates manual transmission to improve fueleconomy and driver’s comfort, are seeing increaseduse.
Conventional AMT’s have typically been
hydraulically or pneumatically operated, but they havedrawbacks in that hydraulic or pneumatic actuation
systems involve a lot of components. This large-74-Ball Screw Unit for Automotive Electro-actuation

3.2 Ball screw unit for variable valve system

number of components leads to relatively large friction
which in turn leads to poor efficiency and the inability
to achieve greater thrust; or slow response.
To be able to solve these problems, NTN has
adopted a ball screw driving system for a gearshift
actuator for automated manual transmission oncompact trucks.

Most automotive gasoline engines are four-stroke
engines that feature four steps of operation—intake ofan air-gasoline mixture into the engine, compression
and combustion in the engine, and exhaust from theengine.
The very mechanical component that governs thiscyclic operation is the intake & exhaust valves in theengine. It is possible to improve fuel economy orpower of a given automobile by varying the timing forthe opening/closing of these valves, or their verticalstroke (lift).
Recently, the need to reduce CO2 emissions within
the automotive industry has been mounting in order tohelp improve the global environment. In this context,
development activities are in progress in theautomotive industry for a stepless-variable valve
mechanism that is capable of more delicate valve
control. In the EU market, marketing of a worm gearbased mechanism for this purpose began several
years ago. Opposite the worm-gear based valve
control system there is the ball screw design, thatoffers low friction and high efficiency. NTN has
developed a unique ball screw unit that boasts a
compact mechanism, higher reliability (safety) andquick response.

3.1.1 Appearance of AMT
Fig. 5 shows the typical appearance of an AMT thatincorporates a gear shift actuator. A gearshift actuator
is mounted on top of the transmission that is originallydesigned for manual gearshift operation to automate
the transmission without much modification to thestructure of transmission.

SELECT side ball screw

SHIFT side ball screw

3.2.1 Structure of variable valve system
Fig. 7 shows the layout of a variable valve system
NTN has adopted.
On this mechanism, rotary motion on the driveshaft
having an eccentric cam is translated into verticalmotion on the valve lifter via a plurality of links (link A,
rocker arm, link B, and output cam).
During this course, the relation between the phaseof the actuator and the control shaft coupled to theactuator is changed to altar the positional relation

Fig. 5 Appearance of AMT

3.1.2 Features of ball screw for AMT
The ball screw unit adopted for this mechanism has
the functionality and durability sufficient enough forautomotive applications. Fig. 6 shows the appearanceof the newly developed ball screws.
The ball screw converts rotary motion into linear
motion at higher efficiency, and develops higher
thrust. For this reason, an AMT can be used while
maintaining the existing basic structure that was
originally intended for manual gearshift operation.

Actuator

Ball screw unit

SHIFT side
Rocker arm
Control shaft

Output cam

SELECT side

Link B

Link A
Driveshaft

Valve lifter

Valve* Courtesy by Hitachi Automotive System

Fig. 6 Ball screw for AMT

Fig. 7 Appearance of variable valve system-75-

NTN TECHNICAL REVIEW No.79（2011）

among the links to invariably control the lift andoperating angle of a valve in the valve system.
When mounted to an automobile, this variable
valve system helps improve output and response of
the engine: consequently, improvements will beachieved not only in fuel economy but also incleanness of emissions. In particular, fuel economy issaid to be improved by approximately 10%.

proceeded with unitization that covers auxiliaries anddeveloped an automotive actuator unit that boasts
quicker response and greater thrust.

3.3.1 Structure of actuator unit
Fig. 9 shows the appearance of NTN's actuator unit,
and Fig. 10 schematically illustrates a cutaway of thisactuator unit.
According to the design specification of the actuator
unit, the efficiency for converting forward/reverse
rotary motion into linear motion with the ball screw has
to be 90% or higher.
The operating shaft is supported by a liner ballbearing in a rolling contact mode; thereby friction on
the screw shaft is low and the ball screw runs
smoothly.

3.2.2 Features of ball screw unit for variable
valve system
The ball screw unit adopted for the variable valve
system has sufficient durability for withstanding
loading from the valve.
In particular, to axially locate the bearing on the ballscrew shaft, a lock nut and retaining ring are usuallyused. In contrast, a bearing is fixed to NTN's newly
developed screw shaft by a staking technique withwhich a part of the shaft is allowed to develop plastic
deformation by staking to lock the bearing. The force
needed to extract the bearing, which has been fixed tothe ball screw shaft by staking technique, is at least 10
times as great as the maximum axial load acting on
the ball screw shaft.

3.3.2 Features of actuator unit
(1) Higher thrustHigher efficiency with the ball screw helps theactuator unit to develop higher thrust.
Fig. 11 shows thrust measurements with NTN’s
newly developed actuator unit. NTN’s actuator unit
develops thrust nearly equivalent to that found via
theoretical calculations (1200 N): also, the actuator
unit has been earning a good reputation from its
users. Note that a sliding screw unit of the same size
develops thrust as low as approximately 1/3 that of
NTN’s actuator unit because of greater frictional forces
occurring on it.

3.3 Actuator unit
To be able to further reduce the size and weight ofthe previously mentioned ball screw unit, NTN has

SpacerStaking

Fig. 8 Appearance of bearing staking

Oil seal
Rubber bootFig. 9 Ball screw actuator unit

Linear ball bearing
Oil seal

Actuating shaft
Shaft cap

Linking member
Synchronizing arm

DC motor

Double-row angular ball bearing
Lock nut
Aluminum case
Ball screw nut

Ball screw shaft

Fig. 10 Structure of actuator unit

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Involute splines

Ball Screw Unit for Automotive Electro-actuation

Generated axial force N

3000

4. Conclusion

2500

This paper has thus described NTN’s newly
developed electromechanical ball screw unit that is
applied to electromechanical automotive mechanisms.
In electromechanical applications for automotive
components, there have been examples of adoption ofa planetary roller screw as a novel mechanism.
Nevertheless, because ball screw units consist of asmaller number of components and feature simplicity
in structure, car designers will increasingly consider
adoption of ball screw drive systems in their designs.
In addition to development of ball screw materials ofhigher reliability and improvements in heat treatment
and machining techniques, NTN is also committed todevelopment for further improvement in efficiency ofball screws, which is an outstanding advantage of theball screw actuator.
At the same time, NTN is attempting to realize
higher functionality, a reduction in size and weight withthe actuator unit, and will remain committed to thefurther expansion of the scope of applicability of its
actuator unit products.

2000
1500

1200N
1000
500
0
3.5

4

4.5
Time sec5
Slider screw

Fig. 11 Thrust force measurements(2) High reliability
For proving immunity to severe environment, NTN's
actuator unit has been subjected to durability tests
under a specified load within high-temperature
atmosphere. To assess performance of the actuator
unit having undergone durability test, the actuator unit
has been further subjected to response testing by
checking operating time measurements.
Even after testing, the actuator unit has maintained
its high response capability: no deterioration is foundeven after durability testing as can be seen in Fig. 12.
Aiming principally at shift-by-wire control system,
NTN has developed a novel actuator unit that boasts
compact structure with a built-in position sensor, andhas already been marketing this novel product. Fig.
13 shows appearance of this actuator unit.

Reference
1) Tateishi, K., et al., "Ballscrew for AutomatedManual Transmission," NTN Technical Review, No.
73, pp. 72-75, 2005.
2) Kazuno, K., "Ball-screw Unit for Variable Valve
Event and Lift System," NTN Technical Review, No.
75, pp. 72-77, 2007.
3) Ikeda, Y., et al., "Electric Ball Screw Actuators for
Automobiles," NTN Technical Review, No. 77, pp.
45-48, 2009.
4) Hamamura, Y., et al., "Developing a New 2.0L
Gasoline Engine with a Continuous Variable Valve
Train" Transactions of the 2007 (Autumn) Congressof the Society of Automotive Engineers of Japan,
pp. 17-22, 2007.

20
15

0.16 sec.

Location mm

10

After
durability test

5
0
0.0
-5

0.1

-10

Beforedurability test

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0

-15
-20

Time sec

Fig. 12 Operating time measurements

Photo of author

Koji TATEISHI
Automotive Business HQ
Axle Unit Engineering Dept.

Fig. 13 Shift-by-wire electromechanical actuator unit

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NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Low-torque Deep-groove Ball Bearings for Transmissions

Katsuaki SASAKI*

To achieve low fuel consumption in response to environmental
concerns, we have focused on reducing the friction of tapered roller
bearings.
NTN has expanded to focus on ball bearings by analyzing friction in
oil lubrication and developing low-torque deep-groove ball bearings fortransmissions. This paper introduces the structure and performance of
these bearings.

The running torque on deep groove ball bearings, inwhich rolling elements roll on the raceway surface inpoint contact, is fairly low. Therefore, further reduction
in running torque with this bearing type was believed
to be near impossible. However, NTN has succeededin reducing running torque on this bearing type by
optimizing the shape of the cage.
Fig. 1 shows NTN’s low torque deep groove ballbearing. This bearing type does not require installationof new production equipment.

1. Introduction
In addressing global environmental issues, the needfor reduced running torque has been increasing fortransmission bearings in order to help improve fueleconomy for automobiles. As discussed in “Market
and Technology Trends in the Automotive Industry” inthis issue of the NTN Technical Review, there havelong been active development efforts about tapered
roller bearings, in which rolling elements are in line
contact with the raceway surface, to reduce friction
loss on automotive transmissions.

2. Low Torque Deep Groove Ball Bearing
2.1 Structure and features
NTN's newly developed low torque deep groove ballbearing is unique in that the contact area between balland cage is decreased by providing a recess in the
pocket of the cage (guide surface for guiding balls) as
shown in Fig. 2 in order to reduce oil shear occurringin this area.
Consequently, the newly developed deep groove
bearing boasts approximately 25% torque reduction
compared with the conventional product.

2.2 About torque-generating factors
As summarized in Fig. 3, there are six factors thatcan be responsible for torque generated on deep
groove ball bearing.
Factors (1) through (4) are governed by the
specification for internal design of bearing and theirmagnitudes can be determined by calculation.

Fig. 1 Low torque deep groove ball bearing* Automotive Business HQ. Automotive Engineering Dept.
-78-

Low-torque Deep-groove Ball Bearings for Transmissions

Magnitude of factor (5) "Shear torque on oil between
cage and balls" is determined by the specification for
cage, while magnitude of factor (6) "Agitation torqueowing to viscosity of lubricating oil" is much affectedby physical properties and amount supplied of the
lubricating oil used.
In recent years, to help reduce fuel consumption of
automobile, oil flow rates in automotive transmissions
has been decreasing. Automotive transmissions havebeen increasingly lubricated with oil mist or near-splash lubrication technique rather than oil bathsystem. Focusing on factor (5) that accounts for alarger percentage in torque occurred under these
operating conditions with NTN’s deep groove ballbearing products, NTN has attempted to develop low
torque variant of the cage.

2.3 Technique for reducing torqueFig. 4 illustrates share breakdown by torquegenerating factors, where this data has been obtainedby running NTN’s deep groove ball bearing 6207
under high speed, low load conditions. Underlubrication conditions that assumes mist or splash
lubrication, torque resulting from “agitation resistancewith oil” is very small. Therefore, NTN has ignored thisfactor from its consideration about torque reduction.
Factors “rolling viscosity”, “differential slipping” and“elastic hysteresis loss on balls” are governed by theinternal design specifications of bearing and bearing
operating conditions. In other words, these factors canaffect life and rigidity of the bearing.
By focusing on “shear resistance on lubricating oil incage” that accounts for approximately 70% of share in
torque generation, NTN has successfully reduced
torque on deep groove ball bearing throughimprovement in cage design.
As can be understood from Fig. 5 and formula (1),reduction in slip area is effective to reduce shear
resistance on oil in the cage. For this reason, a recess
has been added inside the pockets in cages made ofshaped sheet steel shown in Fig. 6 in order to reduce
slip area. Consequently, the size of the contact area
between the balls and cage has been reduced,
thereby decreasing shear resistance from thelubricating oil.

Low torque cage
(Inside the pocket:
recessed shape)

Conventional

Newly developed

Fig. 2 Structure and features(1) Rolling viscosity torque on rolling surface
Torque occurring from reactive force on the oil film thatexists between the rolling surfaces and the balls.
(2) Elastic deformation torque on balls
Torque loss resulting from repeated elastic deformation on
balls.
(3) Torque from differential slipping(4) Torque from spin slipping
Slip torque resulting from difference in circumferential
velocity between the raceway surface and the balls.

Factors that are governed by internal design of the
bearing as well as bearing operating conditions.
These factors affect bearing life and rigidity.

(5) Shear torque on oil between cage and balls
Torque resulting from shear resistance occurring on the oil
between the ball guide surface and balls in the cage.

Torque from rolling viscosity
Torque from differential slipping(6) Agitation torque from viscosity of lubricating oil
Scrape-up torque and scrape-through torque on oil resulting
from actions of balls and cage.

Torque from elastic deformation
on balls

26%

71%

(1)
(2)
(3)
(4)
(5)
(6)

2%
1%

Torque from shear resistance on
oil in cage

Bearing: 6207, Fr = 500N
Ni = 4000 min-1, ATF 30˚C
Mist or splash lubrication
Improved cage helps decrease shear torque andagitation torque on lubricating oil.
This factor does not affect life and rigidity of bearing.

Fig. 3 Possible torque-developing factors
on deep groove ball bearing

Fig. 4 Share breakdown of effects of torque-generating
factors-79-

NTN TECHNICAL REVIEW No.79（2011）

Shear resistanceF=ηus …………Formula (1)hBall

Test bearing
Ni

Slip area s

Oil mist

Viscosityη

Oil film
thickness hFrStaticpressure table
Tangential force
detecting load cellSurface of cage pocket
Enlarged view on area A

Fig. 7 Torque test rig

A
Oil film on cage pocket

Oil mist
application
After oil mist application,
10 sec.

Rotational torqueFig. 5 Cutaway of cage pocket

After oil mist application,
50 sec.

Conventional
New design
0

50

100

150

200

250

Time T sec.

Fig. 8 Result of torque comparison test on 6207 ball bearing
Fig. 6 Low torque cage (inside of pocket)

2.4 Performance assessment

2.4.2 Strength of cage

The performance of unsealed low-torque ballbearings has been tested with NTN’s bearing type6207 which is a typical bearing for automotivetransmissions.

For mechanical strength of the cage, the newly
developed design and conventional design in Fig. 9
were compared by FEM analysis and strength testing.
･ Test bearing: 6207
･ Bearing dimensions:φ35×φ72×17

2.4.1 Result of torque measurement
Bearing samples were tested under the following
test conditions, on the test rig and by the method
summarized in Fig. 7. Fig. 8 graphically shows theresult of the torque comparison test with the 6207
bearing. Due to the effect of the recess in the cage,
the newly developed design shows reduced shear
resistance on the oil film. The torque reduction
reached as much as 25% compared with conventional
design. Also, it has been verified that reduction in areaof the cage pocket corresponds with the theoretical
reduction in shear resistance of oil in the bearing.
･ Test bearing: 6207
･ Bearing dimensions: φ35×φ27×17･ Radial load: 500 N･ Bearing speed: 4,000 min-1･ Lubricating oil: ATF, 30˚C, mist lubrication system

Conventional

Newly developed

Fig. 9 Bearings tested-80-Low-torque Deep-groove Ball Bearings for Transmissions(1) Stress analysis

the result of analysis of the cage model. For either
conventional design or newly developed design, thediagrams in the left are stress distributions viewed
from the ball guide surface, and those in the right are
stress distributions viewed from outside the cage.
As a result of stress analysis, it has been learned
that the greatest tensile stress occurs at corner R areain the vicinity of rivet, and the magnitude of stressgenerated is same with both the conventional design
and newly developed design. Note that stress occurs
on the recess in pocket of low torque cage; however,this stress is much smaller compared to that occurringat the corner R area and does not adversely affectmechanical strength of cage.

The conventional cage and the newly developed
cage were subjected to loading under same
conditions. The resulting stresses in these samples
were compared by FEM analysis technique.
･ Ball position (4): Circumferential loading of 49N*1･ Ball position (1) to (5): (other than (4)): restrained*1 Equivalent to loading when 50% load torque from theengine at 2nd speed is acting on the differential-side
bearing on transmission of small-displacement FF car.

Fig. 10 shows a half model of the cage, and Fig. 11(stress distribution diagram) schematically illustrates(2) Strength verification test

(1)

The most frequent cause of damaged cage isexcessive loading on the cage that results from delay
in travel of the balls when an excessively large
moment acts on the balls. To simulate this situation,
excessive load was exerted on the bearing, thereby
testing cage strength. Table 1 summarizes teh test
conditions for testing mechanical strength of the newly
developed cage and conventional cage, while Fig. 12graphically shows the test result.
Accumulated failure time and mode of failure are
identical in both the newly developed cage and theconventional cage. The failure that occurred was
fracture that started at corner R area and staked pointon the rivet shown in Fig. 13, and correlates with theresult of FEM analysis in Fig. 11. In summary, it has
been positively verified that the cage of the newly
developed deep groove bearing, featuring a recessed
cage, has equivalent mechanical strength to theconventional product.

(2)

(3)

(4)
(5)
Fig. 10 Cage analysis model(MPa)(1) (1)
(2)

90

(2)

60
30

Conventional

(3)

(3)
Conventional0

Table 1 Test conditions-30

Staked point on rivet

(4)

(4)

Corner R

-60

49N

49N

-90

(5)

Stress distribution viewed
from ball guide surface

Momental load

13.7 Nm

Bearing speed

1800 min-1

Run time

Until failure occurs.

(5)
Stress distribution viewed
from outside the cage(1) (1)

(2)

600

(2)

(3)

(4)

Newly
developed
Staked point on rivet

49N

Run time sec

500

(3)

(4)

Corner R

300
200
100

49N

(5)

400

0

(5)

Conventional design

Fig. 11 Result from stress analysis

New design

Fig. 12 Result of cage strength test-81-

NTN TECHNICAL REVIEW No.79（2011）

3. Conclusion

RecessThis paper has described the NTN’s new low torquedeep groove ball bearing technology for automotivetransmissions. This bearing technology boasts
significantly reduced torque without loss in bearing lifeand rigidity. Regarding bearing production, this newtechnology helps realize production of improved
bearing product without the need for introduction of
new purpose-specific production machines. As such,
NTN will market this new bearing as a major bearing
product series.

Staked point on rivet

Corner R

Fig. 13 Names of areas on cage

Reference
1) "Agitation torque that occurs with bath lubrication ofrolling bearings" (reports 1 and 2), 2001 Japan Society
for Precision Engineering Autumn Meeting Academic
Lecture Proceedings.
2) NTN Technical Review, Vol. 6-1, pp. 784-792, 1957.

Photo of author

Katsuaki SASAKI
Automotive Business HQ
Axle Unit Engineering Dept.

-82-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Article ]Technology Trends in Auto Tensioners

Seiji SATO*
With the trend toward high-performance automobile engines, auto
tensioners have been widely adopted for automobile engines since the1980s in order to improve durability and to reduce noise from timing belts
which are used to drive camshafts. In recent years, demands for further
engine downsizing and improved reliability have become strong, andmajor trends include a change from timing belts to timing chains and theuse of serpentine drives with single accessory drive belts. Furthermore,as concern about global warming and other environmental issues
becomes strong, the development of technologies to reduce fuelconsumption have sped up, and demands that auto tensioners be made more compact, lighter inweight and lower friction have increased.
NTN supplies various types of hydraulic auto tensioners for camshaft drive systems and accessory
drive systems. This article introduces the current technologies used in these auto tensioners.

1. Introduction

2. Engineering Trend in Engine Drive
Systems

On automotive engines, a drive belt and/or chain isused to transmit the rotary motion of crankshaft to thecamshaft and auxiliaries. The auto tensioner keeps
correct tension on drive belt and chain.
Recently, people have been increasingly conscious
about environmental issues such as global warming
and needs have been mounting for a reduction inenvironmental impacts associated with automobiles.
Various nations have imposed regulations for
reduction in fuel consumption with targeted values. To
address this situation, engineers in these nations havebeen developing low fuel consumption technologies
that contribute to the reduction of CO2 emissions andincrease energy saving.
In this context, the auto tensioner needs to assist in
the reduction of size and weight of the engine in orderto help improve the fuel economy of the automobile.
At the same time, the auto tensioner incorporated into
camshaft drive & auxiliaries drive systems has been
increasingly expected to contribute to improvements infuel economy and reduction in friction.
This article describes information about NTN’s auto
tensioner products that are intended to meet current
emissions and energy savings market needs.

The camshaft drive system on automotive engines
has to incorporate a meshed transmission-capabletiming belt or chain system in order to synchronize themotion of engine pistons with the cams that govern air
intake and exhaust emission. Fig. 1 shows typicallayout of a timing belt, while Fig. 2 illustrates a timing
chain. Fig. 3 illustrates change in market shares withdifferent camshaft drive systems.
NTN offers a full lineup of auto tensioner products
for both camshaft drive types.
Previously, chain drive systems have been said todevelop friction loss whose level is greater than that ofbelt drive systems (note that friction loss cansignificantly affect the fuel economy of an automobile).
Nevertheless, as silent chain systems boasting quiet
operation have been more commonly employed, thechain drive has gradually superseded the belt drive.
Consequently, the market share of chain drive
systems for camshaft drives has been increasing.
Historically, the chain drive system was notcommonly used for diesel engines because it operates
using engine oil as lubricating oil which is adversely
affected by deterioration of oil by contamination with* Automotive Business HQ. Automotive Engineering Dept
-83-

NTN TECHNICAL REVIEW No.79（2011）soot. However, mechanical strength of drive chains
have been much improved thanks to the improvementin surface treatment technique; thus, chain drive iscurrently more often applied to automotive diesel
engines.
The auxiliary drive systems drive a lot of automotive
auxiliaries using frictional transmission technology
based on V-ribbed belts, etc. Belt drive systems forthe automobile began as a multi-belt configuration
where multiple drive belts are commonly engaged withthe crankshaft and the drive belts respectively transmit
driving power to assigned auxiliaries. The belt drive
system has further evolved into a serpentine drive
system such as a one illustrated in Fig. 4 wheremultiple auxiliaries are driven by one drive belt. This
allowed the length of entire engine system to be

Camshaft timing
pulley (exhaust)

reduced. In the present-day auto industry in Japan,
serpentine camshaft drive systems are used on
approximately 80% of cars newly manufactured inJapan.
The idling stop system, a system that stops theengine during idle, appears to be a useful fuelconsumption reduction technology. For this system, aunique mechanism has been commercialized where
the engine is restarted by power transmitted from thealternator via the auxiliaries belt rather than activating
the starter. For this system, a variety of different auto
tensioner types are arranged within the layout of theauxiliaries belt, wherein each auto tensioner has tomeet requirements for system functionality. This
arrangement can now be seen in new applications.

(%)
100

Camshaft timing
pulley (air intake)

Others80

Share

Timing chain drive system

Tensionidler pulley

60
40
20

Timing belt drive system

02000

Tensioner(year)

2002

2004

2006

2008

2010

Fig. 3 Transition in market shares of camshaft drive
systems

Crankshaft
timing pulley

AlternatorFig. 1 Typical layout of timing belt

Water pump
Idler
Power
steeringpumpCamshaftsprocket
Chain

Auto tensioner

Chain tensioner

Airconditioner

Crankshaft pulley

Lever guideFig. 4 Typical example of auxiliaries drive belt
Fixed guide
Crankshaft sprocket
Running direction

Fig. 2 Typical layout of timing chain-84-Technology Trends in Auto Tensioners

3. Types and Structures of Auto
Tensioners
3.1 Timing belt auto tensioner
Two scenarios are possible to provide correct
tensioning of the timing belt. In one scenario, theamount of tension is determined when the pulley is
installed. In the other scenario, an auto tensioner thatis capable of tension adjustment is incorporated into
the timing belt system. The types of auto tensioners
can be categorized on the basis of damping structure
into frictional systems and hydraulically operated
systems. NTN’s auto tensioner is a hydraulicallyoperated type which offers the benefit of reduced
friction by reducing the initial belt tension.
Fig. 5 schematically illustrates the basic structure of
NTN’s auto tensioner for timing belts.

Basic type
Separate type

Pin
Retaining ring
Oil seal

Cylinder (aluminum alloy)

Air
Oil

Short type

Pulley-arm unitized type

Ultra-compact type

Built-in pulley typeAir-oil two-phase

Rod
Return spring

Wear ring
Plunger

Two-point
support

Separator
Ball
Valve retainer
Plunger spring

Valve sleeve

Shortened passage
Smaller pressure chamber

Fig. 5 Basic structure of auto tensioner

Fig. 6 Various timing belt auto tensioners

A hydraulically operated auto tensioner has adamping structure that utilizes hydraulic oil prefilled in
the interior of its cylinder. To be able to absorb
expansion of oil at higher temperature and maintaining
an adequate seal, NTN employs air-oil two phasestructures and oil seals to contain hydraulic oil. Thechange in belt length caused by temperature change
and aging is accommodated by the operation of thespring in the cylinder; thereby the tension of the timing
belt can be adjusted accordingly. Thus, even though
the belt will inevitably develop wear and elongation,
virtually no change will occur with the tension setting.
Consequently, after the initial setting, the NTN’s auto
tensioner will maintain appropriate belt tension until
the timing belt reaches the end of its useful life.
Fig. 6 shows NTN’s auto tensioner types intended
for various timing belt types.

NTN’s aluminum cylinder design addresses theneed for lighter weights while the smaller size
requirement is met by NTN’s “short type” featuring
shorter axial dimension and “ultra-compact type”
featuring size reduction in all dimensions.
The “pulley-arm unitized” type is essentially a unit
comprised of a hydraulic damper and a pulley arm,
which will lead to lower cost and space saving. The
built-in pulley type helps achieve space saving
because the hydraulic damper is housed within thepulley; thus this auto tensioner can supersede thefixed type tensioner.

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NTN TECHNICAL REVIEW No.79（2011）

3.2 Timing chain auto tensioner

mechanism as well as a serrated screw type thatutilizes a stepless “no-back” mechanism for controlling
frictional resistance on the screw.
The ring type has a “no-back” mechanism that while
permitting “push-in” of the actuator by a preset amountof axial shift, prevents “push-in” of the actuator insteps. This mechanism is a compact system thatutilizes the elasticity of the ring arranged
concentrically relative to the axial center of auto
tensioner.
The serrated screw type is NTN’s unique “no-back”
mechanism, and can accommodate vibration on the
chain through high-speed ranges. It can also maintainthe smallest possible “push-in” when the engine isshut down regardless of its position. Thereby it
contributes to noise damping at engine startup.
Fig. 8 shows the typical shapes of NTN’s timing
chain auto tensioners.
To create a light-weight auto tensioner design, NTN
has long been attempting to employ an aluminum
cylinder. For the ring type, the ratchet mechanism has
been superseded with a register ring formed with wirematerial to create a “no-back” function and simplify the
structure. Consequently, a new auto tensioner
boasting much reduced size and weight has beensuccessfully developed.
As for the serrated screw type, the technical data of
the conventional serrated screw has beenreconsidered. Next, the effective diameter of thescrew component has been reduced without areduction in load capacity in order to further reduce
the weight of the auto tensioner. Note that thisalteration has helped improve the ability to form thethread on the serrated screw resulting in the
production process being drastically shortened. At thesame time, the amount of material scrapped has beensignificantly reduced because the mass-removal
machining process has been deleted.
To be able to reduce running friction on the timing
chain, it is necessary to accommodate a variation inengine torque during continuous engine operation andat engine startup.

On chain drive system, the drive chain between thecamshaft and crankshaft is guided with various sliding
guides; sliding resistance between the chain andguides develops friction loss that can vary dependingon the level of chain tension. For this reason,
optimization of the auto tensioner for timing chain
systems is more important than that for timing belt
systems.
Optimal design of an auto tensioner is generally
determined through tuning on the intended engine
type. Depending on an engine type, control of thechain tension may be necessary by adopting special
damping mechanisms such as a relief valve
mechanism that regulates the internal pressure of the
auto tensioner.
Fig. 7 illustrates a typical example of measuring
operation for tuning an engine.
Previously, mechanical tensioners were used for
motorcycle engines. Now, hydraulically operated auto
tensioners having built-in compact relief valve
mechanisms are available for high-speed motorcycle
engines. Consequently, many big sport motorcycles
adopt NTN’s hydraulically operated auto tensioners,
and this product enjoys a larger share in Japan’s
motorcycle auxiliaries market.
Incidentally, in optimizing an auto tensioner for achain drive, considerations must be given to the
behavior of the chain at engine startup. At engine
start up, oil supply to the auto tensioner will beinsufficient for a short period of time due to a delay inoil supply from the oil pump during start up.
Consequently, the chain can develop wobbling. To
prevent this, a “no-back” mechanism is needed which
prevents “push-in” of the auto tensioner while theengine is being shut down. To address this problem,
NTN employs a ring that features a simplified ratchet

Gap sensor

TargetLoad cell

Serrated screw typeFig. 7 Example of measuring operation for engine tuning

Ring typeFig. 8 Examples of NTN's timing chain auto tensioner-86-

Technology Trends in Auto Tensioners

The serrated screw type auto tensioner is better ataccommodating this variation. With the serrated screw
type, “push-in” distance for the auto tensioner before
the “no-back” function becomes enabled correspondswith the dimension of the backlash on the screw andtherefore can be set to a much smaller distance. The
serrated screw type is very convenient as this functionis active regardless of position on the serrated screw,
and thereby the “push-in” distance can be invariably
adjusted. Consequently, this auto tensioner typeboasts a low noise emission level. Also, thrustoccurring from the spring load within the tensioner isminimized therefore chain tension and frictional losses
are decreased.
Fig. 9 summarizes the results of the friction
reduction effect of NTN’s auto tensioners on an
automotive engine. It should be understood from thisdiagram that the ratchet type (ring type) does not
generate noise when the engine is started and leads
to greater plunger thrust. In contrast, the serrated

3.0

3.3 Auxiliaries drive belt auto tensioner
Previously, multiple belt type drive belt systems
were commonly used to drive engine auxiliaries on
automobiles. Recently, serpentine drive systems thatdrive all of the associated engine auxiliaries with one
belt have been more commonly adopted in order toreduce the overall length and weight of the engine
auxiliaries system, promote a maintenance-freefeature and simplify the belt mounting procedure.
Two types of auto tensioners for auxiliaries drive
belts are available. One type is a friction-based typethat directly utilizes frictional resistance as a damping
force: the other type is hydraulically operated type thatutilizes an amount of prefilled oil in the auto tensioner
as a hydraulic damper. NTN uses the hydraulicallyoperated type since this type is relatively immune tochanges in the characteristics of the belt possibly
resulting from aging and temperature change as well
as stably maintaining belt tension at a predetermined
level.
Fig. 11 shows NTN’s auxiliaries drive belt auto
tensioner.
NTN offers unitized assemblies that include pulleys
and pulley brackets. By using 3D models, NTN

Starting noise with ratchet
type is at acceptable level.
(amplitude: 1mm or smaller)

2.5

Crankshaft torque Nmscrew type does not emit noise even at low plunger
thrust and the resultant crankshaft drive torque is low.
Additionally, NTN has been committed to weight
reduction by an improvement in the method used for
mounting to the engine. Compared with theconventional method for mounting, NTN’s newly
developed self-mounting auto tensioner is unique inthat the threading is formed on the tensioner body sothat the tensioner is directly fastened to the engine.
Thereby any previously needed mounting screws are
removed which positively contributes to a weight
reduction for the automotive engine.
Fig. 10 shows an example of NTN’s self-mounting
auto tensioner.

Difference in torque: 0.47 Nm

2.0
1.5
1.0

Starting noise with serrated screw typeis at acceptable level.

0.5
0
0

10

20

30

40

50

60

70

80

90 100 110 120

Plunger thrust N

Fig. 9 Result of assessment about reduction incrankshaft drive torqueCamshaft sprocket

PulleyPulley bracket

Auto tensioner

Crankshaft
sprocketFig. 10 Self-mounting auto tensioner

Fig. 11 NTN's auxiliaries drive belt auto tensioner-87-

NTN TECHNICAL REVIEW No.79（2011）

mentioned advantages.
Recently, these “idling stop” (no idling) systems
have been increasingly applied to hybrid electricvehicles and large displacement engines. Fig. 13
illustrates a typical layout of NTN’s hydraulicallyoperated auto tensioner with the addition of thefrictional auto tensioner.
Once an automotive engine has been started,
greater drive torque acts on the auxiliaries drive belt
system. Therefore, to prevent the belt from slippingand improve reliability of the belt drive, a frictional auto
tensioner is added immediately downstream of thestarter and alternator so that the frictional auto
tensioner can operate in conjunction with thehydraulically operated auto tensioner. To cope withthe trend of increased popularity of “idling stop”
systems, NTN has developed a shortened auxiliaries
auto tensioner aimed at a reduction in weight and sizeof the auto tensioner.
Fig. 14 shows NTN’s short type auxiliaries auto

investigates possible interference with engine-related
components as well as FEM analysis. This has helped
NTN achieve lighter, yet still highly rigid auto tensioner
design.
NTN’s hydraulically operated auto tensioner is
unique in that it has a spring function for obtaining theminimum necessary tension for the drive belt which
can be set up independent of the hydraulic damper
function. This feature is advantageous because theauto tensioner damper function can be tuned up
without increasing the minimum belt tension evenwhen it has to develop greater damping force.
Consequently, frictional losses on the auxiliaries drive
belt are mitigated.
Recently, in an effort to improve fuel economy ofautomobiles and to help assist in the push for global
energy conservation, idling stop (no idling) systems in
the form of ISG (Integrated Starter Generator) or BSG(Belt Starter Generators) have been increasingly
commercialized. Fig. 12 shows a typical layout of an
auxiliaries drive belt employing ISG system.
In restarting the previously inactive engine, the ISG
system activates the starter and alternator to transmit

Starter & alternator
Hydraulic auto tensioner

Starter & alternator

Auto tensioner
Frictional
auto tensioner

Water pump

Crank pulleyFig. 13 Layout of NTN's hydraulically operated auto
tensioner with addition of frictional auto tensioner
Crank pulley
Resin spring seatFig. 12 Belt layout of ISG system
Oil seal
Springrotary power to the crankshaft via the auxiliaries belt
to start the engine.
The auto tensioner must generate damping force
that prevents the drive belt from slipping even whensubjected to higher torques seen in the restarting
operation of the engine.
When the engine is started by the starter andalternator, the crankshaft (whose inertia is large) is
rotated resulting in higher tension on the belt. To
address this situation, NTN has adopted its unique
hydraulically operated auto tensioner which is capableof greater damping effects because of the previously

Fig. 14 Short-type auxiliaries drive auto tensioner-88-Technology Trends in Auto Tensioners

tensioner. In this auto tensioner, the oil line to the highpressure chamber has been relocated to below thehigh pressure chamber in order to provide sufficiently
large oil prefill space. At the same time, the spring is
now housed in the casing. Consequently, a 15% size
reduction has been achieved. In addition, a resin
made seat for the spring is used in this new auto
tensioner leading to 10% weight reduction.

5. Conclusion
This article has reported the engineering trend seen
in camshaft drive and auxiliaries drive systems forautomotive engines as well as the functions needed
for auto tensioners that are used together with theabove-mentioned systems. Also shown was NTN’s
engineering commitment to its newly developed auto
tensioners.
The need for better fuel economy in automotiveengines appears to be growing in an effort to help
alleviate possible environmental impacts.
NTN will remain committed to the development of its
auto tensioner products by ensuring that ever
diversifying market needs are meet using
improvements in the entire auto tensioner system
including peripheral components.

4. Future of Auto Tensioner
In the automotive market, concept development has
already begun for a system that places top priority ona reduction of frictional loss intending to further reduce
fuel consumption. As already described, a reduction oftension on belt or chain is effective in reducing
frictional loss. For this reason the auto tensioner
needs to develop the minimum necessary damping
force for stability.
Additionally, it will be necessary to optimize thecomprehensive scope of the system. This will include
the adoption of friction reduction technologies NTN
has been developed for peripheral components suchas pulley bearings and chain guides as well as
incorporating other fuel consumption reduction
technologies that will become more increasingly used,such as no idling system.

Reference
1) Koji, A., et al., The Tribology, No. 10, pp. 28ﾐ30,
October 2009.
2) Kita, H., The Tribology, No. 5, pp. 35-37, May 2005.
3) Kitano, S., et al., The Tribology, No. 10, pp. 24-26,
October 2004.
4) Kitano, S., et al., NTN Technical Review, No. 73, pp.
110-117, 2005.
5) Society of Belt Transmission Engineers, Practical
Designs for Belt Transmissions and Precision
Conveyances, New Edition, Yokendo Co., Ltd., 2006.

Photo of author

Seiji SATO
Automotive Business HQ
Axle Unit Engineering Dept.

-89-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]NTN-SNR Low-torque Strut Bearing for Severe Environments

NTN-SNR
NTN

Gérald MIRABEL*
Susumu YAMAJI**

Sealing reliability is a significant issue for automobile strut
bearings, because they must be protected from severe
environmental conditions, such as muddy water intrusion.
Based on its knowledge and experience, NTN-SNR has
developed a second-generation strut bearing.
This article introduces the features and performance of the
bearing’s floating seal.

1. Introduction

2. Structure of McPherson Suspension
SystemAutomotive strut bearings are used in the upper
section of the McPherson suspension system, andallow the strut to smoothly oscillate when the front
wheels are turned.
Recently, demand for cars has been increasing inregions where automobiles encounter severe muddy
water environments. Therefore, improved sealing
performance is needed for these bearings. However,improved sealing performance is typicallyaccompanied by increased torque on the seal, which
can adversely affect stability in steering. Thus,
avoiding an excessive increase in torque on the seal
poses an engineering challenge.
Recently, NTN has developed a unique sealed strut
bearing that boasts both good sealing performanceand low running torque. The following article describes
this unique bearing.

Fig. 1 illustrates structure of a typical McPherson
suspension, and Fig. 2 shows the structure
surrounding the bearing. This suspension structure isused in about 70% of passenger cars around theworld.
The upper section that includes the strut bearing issituated within the tire housing, and has to function insevere environment where it is directly subjected tosplashes of muddy water from the road surface.
NTN-SNR's strut bearings are designed toaccomodate both car type and the structural
surroundings of the bearing. The main product line
within NTN-SNR's strut bearing products to date has
been the GEN2 strut bearing shown in Fig. 3, which
comprises a press-formed bearing race and resin
case.

Suspension mount

Strut bearing
SpringDamperFig. 1 Structure of McPherson suspension**NTN-SNR ROULEMENT Automotive Equipment Engineering**Automotive Business HQ. Automotive Engineering Dept.

-90-

NTN-SNR Low-torque Strut Bearing for Severe Environmentscomponents; the strut bearing is mounted tosurrounding parts such as the suspension mountand the spring.
･ Floating seal (03)･ Press-formed bearing race (04, 06)･ Retainer and balls (05)･ Grease
The GEN2 strut bearing includes the top cover andspring seat, each made of resin: consequently, when
the spring is loaded, the entire bearing will be
deformed. Fig. 4 shows examples of load distribution
patterns on the strut bearing being subjected to aspring load. The length of the coil spring changes
depending on the length of suspension stroke;
therefore, the magnitude of load acting on the strut
bearing will vary greatly. The resulting non-uniform
circumferential load at the end face of the spring leads
to deformation of the strut bearing proper.
This type of deformation occurring on the GEN2
strut bearing may enlarge the size of the passage inthe labyrinth formed between the resin components(that is, top cover and spring seat) on the strut bearing
as shown in Fig. 5; as a result, water can more easily
reach the interior of the bearing.
If water enters the bearing, the bearing race will
corrode and the suspension spring, whitch oscillates
owing to varying load amplitude, will cause the bearing
race to develop fretting wear (FBE: False Brinelling
Effect). While traveling in a car, or in particular,
parking a car in a garage, the car driver will feel
uncomfortable noise and vibration through the steering
wheel. To address this problem, the strut bearing must
incorporate a structure that blocks water ingress.

Suspension
mount
Strut bearing
Suspension
bellows
SpringFig. 2 Strut bearing and surrounding structure

2nd. Generation
strut bearing listing :Top coverSpring seat
Floating seal
Bearing race
Balls and retainer

Fig. 3 Structure of GEN2 strut bearing with floating seal

3. Structure and Features of GEN2
Strut Bearing
Fig. 3 shows exploded view of structure of GEN2
strut bearing.
･ Resin moldings (01 and 02) contain bearing

Labyrinth

Coil spring extended

Coil spring in neutral state

Coil spring compressed

Fig. 4 Load distribution patterns by state of coil spring-91-

Fig. 5 Labyrinth structure andwater ingress route in strut bearing

NTN TECHNICAL REVIEW No.79（2011）

4. Features of Floating Seal

5. Performance Assessment for
Floating Seal
To meet the diverse requirements of various car
makers, NTN-SNR is executing various performanceassessment tests. In particular, NTN-SNR isassessing the sealing performance of various floating
seal designs on actual suspension assemblies so thatfloating seal samples are tested in situations thatrealistically simulate environments automobiles will
experience. Fig. 7 shows the muddy water resistancetest rig that NTN-SNR has used for this test.
The standard test conditions applied are assummarized below:･ Oscillation angle:±40˚･ Loading: 6490 N (typical)･ Muddy water nozzle: 2 pos.
･ Muddy water flow rate: 3 L/min (per nozzle)･ Run cycle: 8 hours (6 hours of spraying, 2 hours ofdrying)･ Test temperature: room temperatureFigs. 8 through 10 show structures of NTN-SNR’sconventional design, NTN-SNR’s floating seal and acompetitor’s molded lip seal; Fig. 11 shows the resultsof the muddy water resistance test for these seals.
NTN-SNR’s floating seal design (Fig. 9) boasts
excellent prevention of water ingress (as much as90%) into the bearing compared with the conventional
design (Fig. 8).
Note that the competitor’s spring cover structure
shown in Fig. 10, having a molded seal lip, boasts
excellent sealing performance during the early stageof use. NTN-SNR has demonstrated that muddy water

Minute shift

Minute shift

Strut bearings used in severe muddy water
environments are directly subjected to large amountsof water and dust in addition to muddy water. To
address this situation, NTN-SNR has recently
developed the unique strut bearing with a floating
seal, shown in Fig. 3. This strut bearing
simultaneously meets both the required sealing
performance and the desired low running torque.
Fig. 6 provides a concept view of the operation of
the floating seal. The seal (03) is in contact with theresin top cover (01) situated above it; thereby it
maintains an appropriate labyrinth in relation to theresin spring seat (02), regardless of its position.
The floating seal can move freely within the slot
formed in the spring seat, independent of radial andaxial displacement between the resin components;
thereby positive contact remains between the floating
seal and top cover. This floating seal not only prevents
ingress of water, dust and muddy water into thebearing but also helps realize low running torque.
The seal material is an important factor for bearing
performance. NTN-SNR employs polyacetal (POM) asa standard seal material for its strut bearing. This isbecause this material not only boasts excellent wear
resistance but also features elasticity that helps tosimplify mounting the seal into the bearing.

Larger gap

Smaller gap

Spring seat (02)shifted to left

Contact pressure
between top cover(01) and seal (03) isalways kept constant.

Strut bearing
Spring seat (02)shifted to rightFig. 6 Contact between floating seal and top cover

Muddy water
nozzleFig. 7 Muddy water test rig for McPherson suspension-92-NTN-SNR Low-torque Strut Bearing for Severe Environments

NTN-SNR’s floating seal

Fig. 8NTN-SNR's conventional design

Fig. 9NTN-SNR's floating seal design

Molded seal lip

Fig. 10Competitor design: molded seal lip

Percentage of samples exhibitingwater ingress %Sealed portion
Bearing portion
Rotational torque

90% reduction in occurenceof water ingress

38% reduction in occurenceof water ingress

NTN-SNR’s
conventional
designNTN-SNR’s NTN-SNR’s Competitor’s
conventional floating
molded seal
designseal design
designFig. 11 Muddy water test results

6. Conclusion
NTN-SNR has developed and standardized thisproprietary seal concept that will positively improvereliability of its strut bearings.
At present, NTN-SNR’s strut bearings with floating
seal have been finding an increasing number ofapplications in newly developed automobiles. NTNSNR hopes to supply its strut bearing products tomore diverse automakers around the world to help
improve reliability of automobiles in travel.

Photo of authors

Susumu YAMAJI

NTN-SNR ROULEMENT
Automotive Equipment
Engineering

Automotive Business HQ
Axle Unit Engineering Dept.

NTN-SNR’s Competitor’s
floating
molded seal
seal design
designFig. 12 Torque test resultssealing performance with this molded seal design is
38% less effective than NTN-SNR’s floating seal
design because the molded seal develops localized
wear owing to unbalanced loading resulting fromdeformation on the resin component.
In addition, Fig. 12 graphically summarizes sharesof contributing factors of seal types to rotational torqueon strut bearing. NTN-SNR’s floating seal design
boasts 50% lower torque versus the competitor’s
design.

Gérald MIRABEL

50% reduction inrotational torque-93-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]NTN-SNR Low-torque High-capacity Pulley Bearing Unit
for Water Pumps

NTN-SNR

Sebastien BRISSON*

Water pumps used on automobile engines are used to circulate the coolant
between the engine and heat exchangers. Water pumps are very importantto the operation of the engine and require a high level of reliability. The seals
and bearings inside the water pump are critical components and require thesame high level of reliability.
As engine technology advances, additional functions of the timing andaccessory belts have increased the load applied to the water pump pulleybearing and also increased the temperature the water pump pulley bearing is
exposed to. The internal design of the bearing must take into account the higher loads and increased
temperatures of today’s water pump pulley bearings.
This paper introduces the structure and performance of a pulley bearing unit developed for water pumps.

durability of the seal, even when the bearing unit isrunning at higher speed, the sealing system needs toallow a minimal amounts of coolant to flow from theseal lip.
In contrast, the newly developed bearing unit forwater pumps (Fig. 2 right) differs from theconventional bearing unit. Two main differences arethat the bearing inner ring is fixed and the bearing
outer ring forms the outer circumference of the
bearing unit; which turns together with the pulley. This
novel structure is realized by the newly developed
flange which helps the bearing outer ring be coupledwith the pump shaft. This flange not only transmits
torque from the outer ring to the pump shaft but also
accommodates possible misalignment with the pump

1. Introduction
There has been an increasing requirement for
reducing CO2 emissions/fuel consumption as well as
improved reliability with industrial products. This trend
is same with a pulley bearing unit for water pumps in
an automotive engine system. To address thesemarket needs, functionality of bearings has to be
improved.
NTN-SNR has recently developed unique, easy-touse, compact, high-performance pulley bearing unit
for water pump (hereinafter referred to as newly
developed bearing unit). The bearing unit employs acoupling system where its outer ring is coupled withthe shaft via a newly developed flange.

2. Structure
Fig. 1 shows an area on the automotive engine
where the water pump is installed. Fig. 2 illustrates
the structure of bearing unit for water pumps. The
primary function of the water pump is to translaterotary motion of drive belt into rotation of the turbine.
On the conventional bearing unit for water pumps(Fig. 2 left), the bearing outer ring is fixed while theshaft (which doubles as the bearing inner ring) canrotate. To prevent outward leakage of coolant, adynamic seal is provided to isolate the bearing interior
from the outside environment. To ensure sufficientFig. 1 Applicable area on automotive engine*NTN-SNR ROULEMENT Automotive Equipment Engineering-94-NTN-SNR Low-torque High-capacity Pulley Bearing Unit for Water Pumps

Pump body
TurbineFlangeDynamic Seal

Shaft
Bearing

Belt contact surface
Structure of conventional water pump

Structure of newly developed water pump

Fig. 2 Structure of bearing unit for water pumpshaft. With a bearing unit size unchanged, this novel
structure helps accommodate a bearing of larger
outside diameter. Because the bearing is larger, thebearing unit can use a larger number of balls andcarry a greater bearing load. Furthermore, the NTNSNR’s newly developed bearing unit allows theloading point from the drive belt to be situated at the
midpoint in bearing width. Consequently, bearing
stress within the bearing is alleviated; thereby thebearing can enjoy longer life. Fig. 3 shows cutaway of
the newly developed bearing unit.
Additionally, as shown in Fig. 4, the outer ring of the
newly developed bearing unit for water pumps can be
fitted with a V-ribbed pulley or timing pulley. In otherwords, a bearing unit of a given size for water pumps
may be used in conjunction with various pulleyshapes.
Fig. 3 Cutaway of newly developed bearing unit

3. Features
There are two operating modes which are possible
with a bearing unit whose outer ring can rotate. In one
mode, a pulley member is simply fitted onto thebearing outer ring like on a bearing for compressors.
In the other mode, the bearing outer ring itself
functions as a pulley member. As with the newly
developed bearing unit, the bearing outer ring is
coupled with the shaft of water pump via a special
flange. The features of this bearing unit can besummarized as follows:
1. With a given bearing unit the size is the same asthat of conventional design. The bearing outer
diameter of the newly developed bearing unit canbe larger; thereby larger balls can be adopted,

V-ribbed pulley

Timing pulleyFig. 4 Pulleys applicable to newly developed bearing unit

-95-

NTN TECHNICAL REVIEW No.79（2011）

helping the bearing unit to be capable of higher load
carrying capacity and longer service life.
2. Possible misalignment from mounting of bearing to
the water pump is reduced.
3. Weight is reduced in the water pump assembly.
4. Standard bearing unit of a given size can beequipped with a wide variety of sizes and types ofpulleys.
Fig. 5 provides a cutaway of coupling between the
bearing outer ring and flange as well as a photo that
illustrates a swaged portion on the outer ring. As canbe viewed from the photo in Fig. 5, the outer rim of theouter ring is fitted over the entire outer circumferenceof the flange by swaging technique. Fig. 6 shows theouter ring swaging process for joining the outer ring to
the flange.
Fig. 7 shows appearance of the NTN-SNR’sbearing unit. The side face of the flange has drilledholes which help simplify installation of the bearing
unit to the water pump casing because the inner ringcan be press-fitted by applying force through theseholes.
Furthermore, to ensure durability, the dynamic seal
described in the previous section needs to allow
minimal amount of coolant away from the seal lip.
Owing to the structure of conventional bearing unit,
coolant flowing from the seal lip can reach the interiorof the bearing; possibly shorten useful life of the
bearing.
In contrast, on the newly developed bearing unit,
coolant is drained away through the drilled holes;
thereby reducing the likely hood of coolant ingress into
the bearing.

4. Performance
4.1 Seal torqueOn the conventional bearing unit, a trace amount of
coolant is allowed to flow from the dynamic seal. For
this purpose, the structure of this seal is very unique,having triplex seal lip. However, this seal structureleads to higher seal torque. Remember that the outerring of the newly developed bearing unit can rotate:
therefore, even when a trace amount of coolant isallowed to flow from the dynamic seal (as in the caseof conventional bearing design described previously),
bearing life is not adversely affected. Thus, a standard
seal such as one having cross-sectional profile in Fig.
8 may be used, leading to lower seal torque. Resultsof bearing torque measurement are given below:･ Entire water pump: 0.05 N . m･ Bearing unit alone: 0.03 N . m

Swaging on outer ring

Fig. 5 Coupling between outer ring and flange

4.2 Coupling torqueOn the newly developed bearing unit, the addition ofa newly developed flange allows the outer ring torotate. To verify that there is sufficiently high couplingstrength between the flange and shaft and outer ring,
the following assessment tests have been performed.

Fig. 6 Swaging process for outer ring

Drilled holeFig. 7 Appearance of newly developed bearing unit

Fig. 8 Cross-sectional profile of standard seal-96-

NTN-SNR Low-torque High-capacity Pulley Bearing Unit for Water Pumps

specification) is 20 N . m for the coupling between theouter ring and flange on the bearing unit for water
pumps. The measurements of coupling torque on
mass-produced bearing units are at around 80 N . m. In
other words, in spite of possible dimensional variation,
the mass-produced products positively satisfy thedesign requirement of 20 N . m.

4.2.1 Coupling torque between shaft and flange
Fig. 9 graphically shows relationship between
interference and coupling torque between the pumpshaft and bearing flange. The coupling torquebetween the shaft and flange is governed by theinterference between these members: appropriate
adjustment of the interference helps maintain optimal
coupling torque. When the interference between theshaft and flange falls in a range of 0.035 to 0.085 mm,
variation in the coupling torque falls in a range of ±10
N . m.
When the interference between the shaft and flange
falls in a smaller range of 0.025 to 0.075 mm, thecoupling torque can vary by 17 N . m at maximum.
NTN-SNR controls the interference on mass-produced
products of its newly developed bearing unit design sothat the coupling torque with mass-produced products
falls in a range of 42 to 51 N . m. More specifically, themass-produced bearing units meet the design
requirement for coupling torque which is 20 N . m or
greater.

4.3 Durability test
The newly developed bearing unit has been
subjected to a durability test under the following test
conditions:(1) High-speed durability test conditions･ Bearing speed: 7500 min-1･ Radial load: 2000 N･ Ambient temperature: 110˚C±10˚C･ Test duration: 1000 h(2) High-load durability test conditions
Bearing speed: 2250 min-1
Radial load: 3500 N
Ambient temperature: 110˚C±10˚C
Test duration: 1000 h
No bearing failure occurred during a series of tests
within NTN-SNR under the above-mentioned test
conditions. Also, no failure has occurred in thecustomer’s bench test.

4.2.2 Coupling torque between outer ring and flange
Fig. 10 graphically plots the result of actual
measurements of coupling torque between the outerring and flange on mass-produced bearing units. Notethat the minimum coupling torque requirement (design

5. Conclusion
This article has presented NTN-SNR’s newly
developed pulley bearing unit for automotive water
pumps, which features a novel bearing structure witha rotating outer ring. In cooperation with the water
pump manufacturer, NTN-SNR will further expand thescope of bearing users, in particular, automakers, andwill contribute to reduction in CO2 emissions andimprovement in fuel economy.

60

Coupling torque (Nm)

50
〜10Nm
〜17Nm

Interference: 0.035 to 0.085 mm

40Interference: 0.025 to 0.075 mm
30
20Minimum required torque

Satisfactory
10
0
0.000

0.010

0.020

0.030

0.040

0.050

0.060

0.070

0.080

0.090

0.100

Interference between shaft and flange

Fig. 9 Relationship between interference and coupling
torque between shaft and flange

Photo of author

Coupling torque (Nm)

Product 1
Product 2
Minimum required torque
100
80
60

Satisfactory40
20

Sebastien BRISSON

0
1

5

9

13

17

21 25 29

33 37 41 45 49 53 57

NTN-SNR ROULEMENT
Automotive Equipment
Engineering

Fig. 10 Creep torque between outer ring and flange-97-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Article ]Development of a System
for Rolling Bearing Design Optimization

Daisuke IMADA**
Tsuyoshi NIWA**
Takashi UENO**
Tomohisa UOZUMI**

There are demands to improve the torque, size, weight, life andrigidity of rolling bearings used in automotive transmissions anddifferentials.
In order to meet those demands, a system for rolling bearing
design optimization has been developed. This paper introduces
this system with application examples.

1. Introduction
People have been increasingly conscious about the
global environment, and, car users are demanding
improved fuel economy for their cars. Needs havebeen mounting for reduced running torque, size andweight for rolling bearings used in automotivetransmissions and differentials (such bearings may behereunder simply referred to as “bearings”). Bearing
manufacturers have been committed to improvements
to bearing design techniques for this purpose 1).
Bearings used in transmissions and differentials
need to meet various requirements including bearing
life, rigidity and safety factors. For improved
automobile fuel economy, bearings need to satisfy
these requirements while also achieving lower torqueand decreased size and weight 2). To develop abearing design that satisfies all of the required
characteristics, a sequence of trial and error needs tobe repeated; in each sequence, a bearing design isdeveloped and then checked for whether the design
satisfies each requirement. This procedure takes a lotof man-hours. There may be cases whererequirements specification is not satisfied under aspecified condition (size). In particular, tapered roller
bearings and angular ball bearings involve many

design factors that can affect required characteristics.
Consequently, it is very difficult to establish a bearing
design that satisfies all required characteristics undera combination of different loading conditions.
In recent years, optimization analysis has been
more commonly used in the field of CAE 3). When
applying optimization analysis to the previously
mentioned bearing design problem, various required
characteristics are set as constraint conditions andcharacteristics including minimum torque, minimum
size, long life and high rigidity are set as objectivefunctions.
NTN has developed a unique bearing design
optimization system that combines a design variables
optimization routine for improving bearing torque or life
and a detailed design routine for realizing thefunctionality of the design variables optimization
routine. This unique design optimization system ishereunder described.
NTN has used a unique design optimization system
for tapered roller bearings commercially since autumn
of 2008. NTN believes this is a pioneering example ofsimilar systems that are currently used throughout theworld.

** Automotive Business HQ. CAE Dept.
** Automotive Business HQ. Automotive Engineering Dept.

-98-Development of a System for Rolling Bearing Design Optimization

design optimization system can be used to optimize
the design of ball bearings (deep grooved ballbearings and angular ball bearings) and tapered roller
bearings. Table 1 summarizes the bearing
combinations whose design can be evaluated with thissystem. The combinations in Table 1 account for
approximately 90% of automotive transmission
bearings.

2. NTN's Bearing Design Optimization
System
2.1 Scope of applicability
Fig. 1 shows an example where tapered roller
bearings are used in a rear differential. NTN’s newly
developed bearing design optimization system adopts
a calculation model as illustrated in this example,
wherein two bearings support one shaft; thus, thisoptimization system is capable of automatic design
optimization for two bearings simultaneously. Thanks
to this capability, a combination of two bearings eachhaving a contact angle (such as tapered roller bearing
and angular ball bearing) may be optimized whileconsidering the balance between the contact angles of
the two bearings. NTN’s newly developed bearing

2.2 Description of functions of design
optimization system
Fig. 2 summarizes calculation process flow forbearing design optimization with NTN’s bearing design
optimization system. First, the operator determines theintended bearing combination, and then inputs thebearing bore diameter, outside diameter, width andloading conditions and other factors, and next sets
variables (such as number of rolling elements); then,the operator executes calculation. The past
performance record-based database may be referred
to for data about existing rolling elements and cages
to help unify bearing designs.
Detail of the calculation sequence can be describedas follows. First, bearing specifications are determinedon the basis of initial values, and bearing life andbearing stress, etc. are calculated and the calculation
result is analyzed. Next, the NTN’s optimization
program reads in the result of analysis: then, whileconsidering the objective functions and constraint
conditions, the program defines variables that lead toa more relevant solution to meet the requirements
specification; and based on the so-obtained technicaldata, the program repeats calculation until optimal
solution is reached. For searching the best solution,
response surface methodology 4) is used. Additionally,
detailed rigidity data about housing, if available, canhelp realize bearing design that reflects housing
rigidity.

Differential
Differential
side-bearing

Pinion shaft
support bearing

Rear side
Front side

Fig. 1 Typical application to tapered roller bearing
Table 1 Possible combinations of bearings
Deep grooveball bearing

Angularball bearing

Taperedroller bearing

Deep grooveball bearing

○

○

−

Angularball bearing

○

○

−

Taperedroller bearing

−

−

○

Optimal calculation

Alteration to database,reflecting degree of influence
to objective functions

Customer’s
requirements• Rollingelements• CageNO

Selection of
optimal combination
Deep grooved
angular ball bearing

Tapered roller bearing

Bearing data setup

Alteration to database, reflecting
degree of influence to objectivefunctions

Calculation

Calculation program
Consideration of housing
rigidity, bearing life,
bearing stress, etc.

Constraint conditions
for objective functions
Feedback
YES

Result of analysis

OutputOptimal solution
for each set of specifications(Longer life, lower torque, lighter weight, higher rigidity)

Fig. 2 Calculation process on NTN's bearing design optimization system-99-

NTN TECHNICAL REVIEW No.79（2011）

Typical examples of a “calculation program” asfound at the upper right corner of Fig. 2 are as follows:[1] Program for determining bearing load on the basisof load acting on the shaft system.
[2] Program for calculating loading on rolling elementsof the bearing and contact bearing stress on rollingsurface.
[3] Program for calculating dimensions of relevantcrowning.
[4] Program for calculating bearing load capacity.
[5] Program for calculating bearing torque.
[6] Bearing life calculation program.
[7] Bearing mass calculation program.

3. Examples of Design Study Activities
3.1 Overview of optimization
Using this newly developed bearing design
optimization system, NTN has studied optimization oftapered roller bearing, with lower torque as theobjective function, under constraint conditions aimed
at reducing the bearing width. This design study is
described below.
The subject of this design study was the pinion shaft
support bearings for the automotive differential as
shown in Fig. 1. Fig. 4 illustrates the results of NTN’s
bearing optimization study.
Before the design study, the bearing design
satisfies life requirement with a certain margin. The
design of the bearing internal geometry is determined,
wherein lower torque is taken as the objective function
and the life requirement is the constraint in order to
optimize the bearing design so to satisfy the life
requirements and minimize the bearing torque.
Through this design optimization, the resultant

NTN's bearing design optimization system is capableof the following settings for objective functions:･ Calculated life
･ Torque･ Mass･ Bearing rigidity
The bearing design optimization system is also
capable of the following settings for constraint
conditions, as well as settings for upper and lower limits:･ Calculated life
･ Safety factor･ Bearing stress･ Bearing rigidity･ Torque･ PV value on rib (only for tapered roller bearing)･ Bearing dimensions (design dimensions); etc.

Before design study

2.3 Effects of bearing design optimization system

Front side

By adoption of this unique system, NTN has
succeeded in significantly reducing the time of thedesign study phase compared with the design
conventional system.
Fig. 3 shows a comparison in terms of days neededfor design (design study only). The effect of thissystem is particularly apparent for shortening thedesign phase associated with tapered roller bearings.

Optimization

After design study

Torquereduction

Rear side

Front side
Beforedesign study

System not applied
System applied

Design time required

1/8

After
design study<<Bearing life>>

1

1/6

Rear side

1/10

Life
requirement

0.8
0.6

Front side

0.4<<Torque>>Rear side

Torque reduction

Torque reduction

0.2
0Deep grooveball bearing

Angular contact
ball bearing

Taperedroller bearing

Front side

Fig. 3 Comparison in terms of design study time

Rear side

Fig. 4 Optimization activities for pinion shaft support bearings-100-Development of a System for Rolling Bearing Design Optimization<<Example of optimization study 1>>
Optimization study for maximum bearing life

bearing design will satisfy functional requirements
including life and dimensions, while featuring
minimized torque.

<Objective function>
3.2 Typical example of optimization

Life: maximized<Constraint conditions>

Another example of a design optimization study fortapered roller bearings using this system is described
below. The bearings in this design study are tapered
roller bearings for the differential on a CVT as shownin Fig. 5.

Torque: Not greater than conventional bearing
designBearing stress: Not greater than conventional
bearing design
Bearing dimensions: Same as conventional bearing
designTapered roller bearings for differential side on CVTUsing the test bearings, optimization to maximize
bearing life has been executed. Fig. 6 summarizes theresults of comparison in terms of bearing
specifications and bearing performance.
Generally, a increasing the contact angle (a) of atapered roller bearing leads to a drop in the load
rating, while bearing stress and torque also decrease.
This is because the engine-side bearing in particularcarries greater stress load, and an increased contact
angle leads to improvement in bearing thrust load
capacity.
The present bearing design optimization was
intended to attain balance between the contact anglesof the right-and left-bearings, which helps maximize
the calculated bearing life; at the same time, thebearing stress and torque have been reduced.

Enlargedview

Engine-side bearing
Opposite-to-engine-side
bearing

Fig. 5 Structure of differential on CVT

Bearing specification

Bearing performance

Opposite-toengine side

Engine side

1.10

(Performance is represented as total
ratio of engine-side torque with opposite-to-engine-side torque)
Torque ratio
Life ratio

Beforedesign study

1.05

Contact angle α=22˚Cr =54,000 N
Cor =65,000 N

Contact angle α=15˚Cr =52,500 N
Cor =65,500 N

Bearing stress ratio

1.00

0.95

After design
study

α

0.90
Contact angle α=23˚15'
Cr =53,000 N
Cor =70,000 N

Contact angle α=15˚22'
Cr =54,000 N
Cor =67,000 N

Fig. 6 Optimization result 1-101-

Beforedesign study

Afterdesign study

NTN TECHNICAL REVIEW No.79（2011）<<Example of optimization study 2>>
Optimization study for minimum torque4. Conclusion<Objective function>

The bearing design optimization system describedabove boasts the following features:

Torque: minimized<Features of rolling bearing design optimization system><Constraint conditions>• Bearing design can maximize or minimize objectivefunctions (life, torque, mass, bearing rigidity) while
meeting constraint conditions.
• For bearings with contact angles such as tapered
roller bearings and angular contact ball bearings,
which are used in pairs to support a single shaft, thebearings can be designed so that good balance isachieved between the contact angles of right andleft-bearings.
• Unification of bearing types is possible by referring to
the database of past designs for rolling elements,
cages and other bearing components.
• Time required for design study has been reduced
compared with conventional design system.
• Less experienced designers can design bearing withcompetence of experienced designers.

Life: Greater than conventional bearing design
Bearing stress: Not greater than conventional
bearing design
Bearing dimensions: Same as conventional bearing
designUsing the same test bearings, optimization tominimize bearing torque has been executed. Fig. 7
summarizes the results of comparison in terms ofbearing specifications and bearing performance.
In this optimization study, increased contact angle
and shorter roller length were selected in order toreduce bearing torque, resulting in a smaller load
rating. Nevertheless, because the balance between
contact angles of the right and left bearings has been
optimized, the calculated life of the optimized bearing
system is equivalent to that of the conventional
design, and bearing stress on optimized bearings issmaller.

Use of this bearing design optimization system
helps shorten the time required to design bearings
optimized for requirements such as reduction in
torque, or reduction in size and weight. Consequently,
NTN can now more promptly present optimal bearing
designs to its customers.

Bearing specification
Engine side

Bearing performance

Opposite-toengine side

1.10

(Performance is represented as total
ratio of engine-side torque with opposite-to-engine-side torque)
Torque ratio
Life ratio

Beforedesign study

Contact angle α=22˚Cr =54,000 N
Cor =65,000 N

1.05

Contact angle α=15˚Cr =52,500 N
Cor =65,500 N

Bearing stress ratio

1.00

0.95

After design
study
0.90
Contact angle α=25˚57'
Cr =48,000 N
Cor =58,000 N

Contact angle α=16˚1'
Cr =51,000 N
Cor =64,500 N

Fig. 7 Optimization result 2-102-

Beforedesign study

Afterdesign study

Development of a System for Rolling Bearing Design Optimization

Reference
1) Tsujimoto, T., Mochizuki, J., "High Capacity Tapered
Roller Bearings-Super Low Torque, High Rigidity
Tapered Roller Bearings," NTN Technical Review, No.
73, 2005.
2) Kanamoto, T., Ueno, T., Katayama, A., Satou, M.,"Transmission Technology Trends and ProductDevelopments," NTN Technical Review, No. 75, 2007.
3) Nagatani, H., Niwa, T., "Application of Topology
Optimization and Shape Optimization for Developmentof Hub-Bearing Lightening," NTN Technical Review,No. 73, 2005.
4) Yamakawa, H., ed., Optimal Design Handbook,
Asakura Publishing Co., Ltd., 2003.

Photo of authors

Daisuke IMADA

Tsuyoshi NIWA

Takashi UENO

Tomohisa UOZUMIAutomotive Business HQ
CAE Dept.

Automotive Business HQ
CAE Dept.

Automotive Business HQ
Axle Unit Engineering Dept.

Automotive Business HQ
Axle Unit Engineering Dept.

-103-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Paper ]Rapid Evaluation of Shear Fatigue Properties of Rolling Bearing Steels
for Lifespans Up to the Gigacycle Range
Noriaki SAKANAKA*
Yukio MATSUBARA*
Yoshinobu SHIMAMURA**
Hitoshi ISHII***In rolling bearings, subsurface flaking failures occur under pure rolling contact
fatigue conditions. In such failures, one of the crack initiation factors is believed to be
repetitive orthogonal shear stress. The shear fatigue properties can be obtained bytorsional fatigue testing. However, up until now, it has been all but impossible toestablish the shear fatigue properties in the gigacycle range because of the limits of
conventional low-loading-frequency torsional fatigue testers. In contrast, we
employed ultrasonic fatigue testing to assure high loading speeds.
We have developed an ultrasonic torsional fatigue tester that enables the rapid evaluation of shear fatigue
properties. The loading frequency of 20,000 Hz is quite high, resulting in a radical reduction of testing duration.
For example, when loading cycles of 1010 are applied, a conventional tester (with a loading frequency of 10 Hz)
requires over 32 years to complete testing. In comparison the ultrasonic torsional fatigue tester requires only 7
days. This tester is effective for evaluating the properties of rolling bearing steels.

One example of a rapid evaluation method for sheer
fatigue properties of rolling bearing steel is ultrasonicaxial load fatigue testing. Many examples of
evaluation results from this technique are available. In
contrast, very little researchof ultrasonic torsional
fatigue test had been reported by the end of 2007,
wherein the materials tested were Al alloy 2–4) and mild
steel 5) and the maximum amplitude of shear stressapplied was approximately 260 MPa. In 2008 andbeyond, two results of evaluation for high strength
steel have been published; one for spring steel of
598HV 6) and the other for carburized steel of 817HV 7).
According to these results, the maximum shear stressamplitude of the former material is approximately 600
MPa, and that of the latter material is approximately
800 MPa.
Conventional ultrasonic torsional fatigue testers
were not capable of rapid analysis of high-strength
steel materials under a wider loading range. To
address this problem, NTN has developed a novel
ultrasonic torsional fatigue tester (loading frequency;
200,000 Hz, alternating) for rapid evaluation of shear
fatigue properties of high-strength rolling bearing
steels under a wider loading range. This paper
describes NTN’s development work with its unique

1. Introduction
One of the stresses that can trigger crack initiation
prior to subsurface flaking failure in rolling bearing isrepetitive orthogonal shear stress (near-purely
alternating) that repeatedly acts on the surface ofbearing1). Shear fatigue characteristics of a given
bearing steel can be determined by torsional fatigue
testing. For tension-compression fatigue testing (axial
load fatigue test and rotating bending test), so far, it
has been a generally accepted practice to take thefatigue strength of steel having undergone 107 loading
cycles as the fatigue limit of the material. However,even under heavy loading conditions, adequately
lubricated rolling bearings do not develop subsurface
flaking failures within 107 loading cycles. Therefore, inorder to determine the shear failure properties of agiven bearing steel, it is necessary to extend theduration of evaluation up to the gigacycle range.
Incidentally, conventional torsional fatigue testers
apply low frequency loading. For example, with aloading frequency of 10 Hz, it will take more than 32
years to apply 1010 loading cycles. This means
evaluation for bearing steel in the gigacycle range isvirtually impossible.

***Advanced Technology R&D Center***Department of Mechanical Engineering, Faculty of Engineering, Shizuoka University***Emeritus Professor of Shizuoka University-104-

Rapid Evaluation of Shear Fatigue Properties of Rolling Bearing Steels for Lifespans Up to the Gigacycle Range

2.2 Design of specimen

ultrasonic torsional fatigue tester and evaluation ofshear fatigue properties of bearing steel JIS-SUJ2.

Fig. 2 schematically shows the test specimen. The
specimen is a dumbbell type that is defined withshoulder length L1, half chord length L2, shoulder
radius R2, minimum radius R1 and arc radius R. In
designing the specimen, arbitrary L2, R2 and R1 are
first chosen, and substituted into Expressions (1)through (6) along with resonant frequency = 20,000
Hz, E = 204 GPa,ν= 0.29,ρ= 7,800 kg/m3 (E,νandρare actual measurements taken on SUJ2) Thereby,a theoretical solution L1 is determined 5). Arc radius R
can be determined with R1, R2 and L2. At first, a test
piece was prepared whose length satisfies thetheoretical solution L1; however, this specimen did not
develop resonance.

2. Ultrasonic Torsional Fatigue Tester
2.1 Constitution of NTN's ultrasonic torsional
fatigue tester
Fig. 1 schematically shows the constitution of thetester. The major components of the tester are a
torsional vibration converter (loading frequency range:
20,000 ± 500 Hz), amplifier and torsional vibration
magnifying horn. Though having a relatively low rated
output of 300 W, the torsional vibration converter iscapable of applying a stable, high shear stress to thespecimen by optimizing the configuration of thespecimen and torsional amplitude magnifying horn.
As a result of repeated application of the shear
stress onto a specimen of rolling bearing steel, a crack
will eventually occur on the specimen. When theseverity of the crack reaches a particular level, theresonance frequency of the specimen will drop. Basedon this knowledge, we have developed a test
procedure; according to which, when the resonancefrequency of the specimen exceeds a preset variation,
the specimen is regarded as failed and the test is
terminated.

E

G＝

2 (1+ν)………………………………（1）ω＝2πf ……………………………………（2）

α＝

1
L2

k＝ω

R22

（ ）

arccos hρGR12…………………（3）…………………………………（4）β＝ α2 - k2 ………………………………（5）
L1＝

1

1
arctan ｛βcoth（βL2）-αtanh（αL2）｝…（6）
k
k

［

］

Then, eigenvalue analysis for free torsional
resonance was executed on the basis of FEM.
Consequently, the resonance frequency of thespecimen FEM model whose theoretical solution was
L1 was 19,025 Hz, which did not fall in the loading
frequency range of the torsional vibration converter(20,000 ± 500 Hz). Through eigenvalue analysis, L1
value that results in a resonant frequency of 20,000
Hz was determined.
Fig. 3 graphically plots the axial distribution of angle
θ and shear stress τ on surface of specimen
determined by eigenvalue analysis on free torsional
resonance on the specimen FEM model. The torsional
angle at the end face of the specimen was assumed
to be θend = 0.01 rad, wherein the maximum shear
stress acting on the surface of minimum diameter
section of the specimen was τmax = 526 MPa.
Consequently, the relationship between θend and τmaxin the linear elasticity region can be shown by
Expression (7) as follows:

Torsional
vibration converter

Specimen
Torsional amplitude
magnifying horn

Fig. 1 Schematic of ultrasonic torsional fatigue testerτmax＝5.26×104θend ……………………（7）-105-NTN TECHNICAL REVIEW No.79（2011）

Incidentally, the specimen in Fig. 2 lacks a parallel
section. Thus, minor stress concentration occurs at
the minimum diameter area. Nevertheless, the stressconcentration coefficient is α= 1.054 8), and is not a
great value. Note that the possible values tmax
determined by Expression (7) include α= 1.054.

Fig. 4 shows the shape of a highly efficient torsional
amplitude magnifying horn as well as graphicalplotting of the axial distribution of the torsional angle
obtained from eigenvalue analysis for free torsional
resonance on the FEM model. The amplitude
magnifying horn was designed and adjusted so that it
develops resonance at around 20,000 Hz. The
specimen material is Ti alloy ( E = 116 GPa,ν= 0.27,
ρ= 4.460 kg/m 3). When the torsional angle at thelarge end face is θend_L = 0.000418 rad, the torsional
angle at the small end face is θend_S = 0.018 rad. Thismeans the torsional angle magnifying power is 43.1.

2.3 Optimized relationship in shape between
specimen and torsional amplitude
magnifying horn
The magnitude of the maximum shear stress acting
on the surface of the specimen’s minimum diameter
section is greater with a larger specimen size.
However, a larger specimen size resulted in unstable
resonance, and unwanted stoppage of the test.
Therefore, the specimen size was determined so thatthe specimen can stably develop resonance up to
90% of the amplifier output and at the same time themagnifying power of the torsional amplitude
magnifying horn was increased.

0
-10
-20

Axial location mm

Minimum
radiusR1

θend_L=-0.000418 rad

Shoulderradius
R2

-30
-40
-50
-60
-70
-80
-90
-0.02 -0.01

Arc radius R

0θ
Shoulder Half chord
lengthlength
L2
L1

FEM model

Fig. 4 Shape of a highly efficient torsional amplitude magnifying
horn and the axial distribution of torsional angle distributionθend=0.01radτmax=526MPa

15

15

10

10

Axial location mm

Axial location mm

0.02(Magnifying power; 43.1)

Fig. 2 Schematic of specimen

5

0

-5

-10

-15
-0.02

0.01

rad
θend_S=-0.018rad

5

0

-5

-10

FEM model

-15
-0.01

0

θ

rad

0.01

0.02

0

200

τ

400

600

MPa

Fig. 3 Axial distribution of angleθand shear stressτon the surface of the specimen determined by eigenvalue analysis on free torsional resonance(When torsional angle θend = 0.01 rad at end face of specimen)-106-

Rapid Evaluation of Shear Fatigue Properties of Rolling Bearing Steels for Lifespans Up to the Gigacycle Range

2.4 Amplitude measurements of maximum
shear stress acting on the surface of thespecimen’s minimum diameter section

established between P and θend as shown in Fig. 6.
Thereby Expression (8) was obtained which
represents a regenerative line. Expression (9) was
developed using (7) and (8), defining the relationshipbetween the amplifier output P and the maximum
torsional stress amplitude τmax acting on the surfaceof the specimen’s minimum diameter section. From
Expression (9), when P = 90%, τmax = 951 MPa.
Thus, it has been found that this amplitude is great
enough to cause high-strength rolling bearing steel todevelop shear fatigue fracture.

Table 1 summarizes chemical components in highcarbon chromium bearing steel JIS-SUJ2 used tofabricate the specimen. The specimen was preparedthrough turning, heat treatment and finish by grinding.
The heat treatment process consisted of heating(830˚C for 80 min.), oil hardening (80˚C), andtempering (180˚C for 180 min.).
Using three pieces of the specimen, the torsional
angle θend (rad) on the end face of the specimen wasmeasured with varied amplifier output P (%). The
amplifier output P was varied from 10% to 90% in 5%
increments Thereby, the torsional width at the end
face 2a was measured and torsional angle θend wasdetermined on the basis of the relationship illustrated
in Fig. 5. Consequently, a linear relationship has been

θend＝（1.96×10-4）P + 4.35×10-4 ………（8）τmax＝5.26×104θend＝10.3 P + 22.9 ……（9）2.5 Effective loading frequency in intermittent
loading mode
Fig. 7 shows the scheme of intermittent loading 9). A
period of 0.01 s is needed before the specified
amplifier output is reached. This period is designatedrise time ts. The number of loading cycles in thisperiod is not included in the count of loading cycles.
Also, when the predetermined loading time td has
elapsed and the amplifier output has dropped to zero,vibration will gradually subside, rather than stopping
immediately: loading cycles during this attenuation
period are not counted, but the time is included in the
non-loading time tp. Therefore, when loading
frequency is taken as f, then effective loading
frequency fe can be determined by Expression (10)below:

Table 1 Chemical components in SUJ tested(in wt%; ppm for O)
C

Si

Mn

P

S

Ni

Cr

Mo

Cu

0

1.02 0.27 0.43 0.014 0.007 0.05 1.48 0.04 0.09

6

θend
2a

fe ＝

Torsional angle at the end face of specimen θend rad

Fig. 5 Relationship between torsional width 2a andtorsional angleθend at the end face of specimen
0.020

f・td
(ts + td + tp)…………………………（10）Rise time ts

Specimen 1(0.01s)Specimen 2

Loading time td

Specimen 3

0.015

Non-loading time tp
Attenuation time

t

0.010
1 process cycle
Number of cycles before rise time and during attenuation period
are not included in count of loading cycles.

0.005

Fig. 7 Scheme of intermittent loading
0.000

0

10

20

30

40

50

60

70

80

90

100

Amplifier output P %Fig. 6 Relationship between amplifier output P andtorsional angle at the end face of specimen θend-107-

NTN TECHNICAL REVIEW No.79（2011）

operation of the ultrasonic torsional fatigue tester. The
solid line in Fig. 8 is an S-N line fitted to the fatigue
limit polygonal line model per metal material fatigue
reliability evaluation standard JSMS-SD-6-02 of The
Society of Materials Science, Japan where the shear
fatigue limit is τw0 = 577 MPa.
Observation of a particular area on the surface ofeach failed specimen, to which the maximum stressacts, reveals that initial shear-mode crack in the axial
direction as shown in Fig. 9 appears on all the
specimens.

3. Evaluation of Shear Fatigue
Properties of SUJ2 by Ultrasonic
Torsional Fatigue Test
Shear fatigue properties of the SUJ specimen—thesame specimen type previously described in Sec. 2.4
was tested. In order to eliminate possible effect of
surface roughness, the nodes on the specimen weremirror-finished by subjecting these nodes to polishingwith emery paper (#500 and then #2000) and lappingwith diamond powder (grain size 1 µm) while turning
the specimen.
To avoid heat buildup on the specimen, dry air was
blown to the nodes for force-cooling. During testing,the temperature on the minimum diameter section wasmeasured with a radiation thermometer; as a result, it
was determined that significant heat is generated at ashear stress amplitude of 700 MPa or greater.
Virtually no heat generation was found at an amplitudeless than 700 MPa. For this reason, one phase of up
to 107 cycles was designed to consist of 0.11 s ofloading and 1.1 s of non-loading. If the specimen did
not fail before the completion of this phase the test
was temporarily suspended, and then resumed with aloading cycle consisting of 1 s of loading and 0.2 s ofnon-loading. If the specimen did not develop failure
before 1010 cycles, the test was terminated. If theloading frequency is set to 20,000 Hz, then by
calculation with Expression (10) given in Sec. 2.5, theeffective loading frequency is 1,803 Hz with eachcycle consisting of 0.11 s of loading and 1.1 s of nonloading, and is 16,529 Hz with each cycle consistingof 1 s of loading and 0.2 s of non-loading.
Fig. 8 graphically plots the shear fatigue properties
of SUJ2 obtained from evaluation by intermittent

4. Discussion
NTN's supersonic torsional fatigue tester(alternating) not only functions as a rapid torsional
fatigue tester but also appears to provide the ability toestimate fatigue limit contact pressure of rolling
bearing steel by evaluating shear fatigue properties up
to giga-cycles range (e.g. 1010 loading cycles). Rolling
bearings have recently been used for greatly
prolonged service periods. To address this situation,
fatigue limit contact pressure has been used in the life
calculation formula in the relevant ISO standard 10).
Practice for estimating fatigue limit contact pressure isdiscussed below. Incidentally, there is a report about
evaluation of fatigue properties of SUJ2 in a manner
similar to that in this paper, with a conventional
torsional fatigue tester. This report states that there is
no shear fatigue limit with SUJ2 because all the
specimens failed within 1010 loading cycles 11). From
Fig. 8, it appears that 108 loading cycles are in thetime-strength regime, and more loading cycles arenecessary to be able to determine the shear fatigue
limit of SUJ2.
From Fig. 8, it is apparent that the shear fatigue
limit of SUJ2 isτw0 = 577 MPa. In line contact
scenario, the maximum contact pressure Pmax is four

Shear stress amplitude MPa

1200
1000

S-N line
800

600
τw0=577MPa

400

200
0
1.E+05

1.E+06

1.E+07

1.E+081.E+091.E+10

Shear-mode
crack

1.E+11

Number of loading cycles

Fig. 8 Shear fatigue properties of SUJ2 evaluated byultrasonic torsional fatigue test

Fig. 9 Magnified view of initial shear-mode crack-108-Rapid Evaluation of Shear Fatigue Properties of Rolling Bearing Steels for Lifespans Up to the Gigacycle Range

effect, where the fatigue limit is lower at the larger
diameter on the parallel section and gradually nears
the fatigue limit found in axial load fatigue test 12). In
NTN’s test method, three different steels each having
unique tensile strengths have been subjected to axial
load fatigue testing and rotary bending fatigue testing(with varied diameters on parallel section) to
determine the fatigue limit for each steel type. As aresult, it has been learned that the fatigue life in axial
load fatigue testing, regardless of the steel type, isapproximately 80% that of the fatigue limit resulting
from rotary bending testing on the parallel section ofdiameter 4 mm.
In summary, the fatigue limit in the axial load fatigue
test, not having a stress gradient, makes the standard
safe in tension-compression fatigue testing. On theother hand, because not having a stress gradient, thetorsional fatigue test does not provide such astandard, even when the diameter of the parallel
section is enlarged as much as possible. As long as
there is a stress gradient, torsional fatigue testing
cannot be free from size effects. To address thisissue, it was assumed in Fig. 10 that the standard fortension-compression fatigue testing can be applied
without any modification to torsional fatigue testing. In
other words, because the minimum diameter of theultrasonic torsional fatigue specimen measures 4 mm,shear fatigue limit τw0 is multiplied by factor 0.8,
resulting in τw0 = 412 MPa.

times as great as the repetitive orthogonal shear
stress amplitude t0, as described in Sec. 4.3. To beable to estimate the fatigue limit contact pressure, we
believe the following two modifications are necessary.
Fig. 10 graphically plots the shear fatigue limit τw0
that reflects the following modifications:(1) Modification for failure probability (see Sec. 4.1)
(2) Modification for size effect (see Sec. 4.2)

4.1 About modification for failure probability
Results of fatigue testing include variation.
Therefore, it is preferable that fatigue properties be
determined through statistical techniques. For thispurpose, a P-S-N line may be developed by evaluating
several specimens at various stress levels. However,
implementing this technique will be difficult in manycases due to limited time. The metal material fatigue
reliability evaluation standard JSMS-SD-6-02 of The
Society of Materials Science, Japan was applied todetermine the shear fatigue limit τw0 from the shear
fatigue properties in Fig. 8,. This standard provides
the ability to develop a P-S-N line from a smaller
number of data sets. The broken line in Fig. 10 shows
the resultant P-S-N line at 10% failure probability,
wherein the shear fatigue limit is τw0 = 515 MPa.

4.2 About modification for size effect
In the torsional fatigue test the specimen exhibits ashear stress gradient that is highest at the surface ofspecimen and lowest at the center of specimen. It his
known that in various tension-compression fatigue
tests, the vertical stress is uniform within the specimen
cross-section at the parallel section, and has aparticular fatigue limit regardless of the diameter. In
contrast, rotary bending fatigue testing has a stressgradient. It is known that the specimen exhibits a size

4.3 Estimation for fatigue limit contact pressure
in line-contact mode
The relationship between the ratio of the major axis
radius (a) to the minor axis radius (b) of contact ellipse(b/a) and 2τ0 /Pmax can be approximately defined byExpression (11) 13). The fatigue limit contact pressure
can be estimated by substituting value of b/a inExpression (11). In the case of line contact mode, thatis, b/a→0, Pmax is 4 times as great asτw0 fromExpression (11):

1200
SUJ2

Shear stress amplitude MPa

S-N line
1000
P-S-N line
(failure probability: 10%)
800

2τ0
Pmax

τw0=577MPa

600-1.12×10-1

4000
1.E+05

τw0=515×0.8=412MPa

1.E+06

1.E+07

1.E+08baba+4.86×10-2

2-7.44×10-3

b
a

3

b
+5.00×10-1a…………………（11）τw0=515MPa
200

4

（） （）
（） （）

＝-1.00×10-3

1.E+091.E+10The shear fatigue limit having undergone
modification for failure probability and size effects isτw0 = 412 MPa. Accordingly, the fatigue limit contactpressure is estimated at Pmax lim = 4τw0 = 1,648 MPa.
Incidentally, the latest ISO standard (2007 revision) 10)
specifies the fatigue limit contact pressure of 1,500

1.E+11

Number of loading cycles

Fig. 10 Shear fatigue limitτw0 with modification for failure
probability and size effect-109-

NTN TECHNICAL REVIEW No.79（2011）

MPa. Our experimentally obtained fatigue limit contactpressure, mentioned above, is virtually same as thevalue in the ISO standard.

Reference
1) Lundberg, G. and Palmgren, A., Acta Polytechnica,
Mech. Eng. Series, 1-3, 5-58, 1947.
2) Stanzl-Tschegg, S., Mayer, H., and Tschegg, E.
Ultrasonic, 31, 275-280, 1993.
3) Iijima, J., Ishii, H., Araki, H., Tohgo, K., High cycle
torsion fatigue characteristic by ultrasonic fatigue
testing method, Proceedings of the 54th Tokai Branch
Regular Meeting of the Japan Society of MechanicalEngineers, 131-132, 2005.
4) Mayer, H., Int. J. Fatigue, 28, 1446-1455, 2006.
5) Garcia, I.M., Doucet, J.P. and Bathias, C., Int. J.
Fatigue, 29, 2094-2101, 2007.
6) Akiniwa, Y., Stanzl-Tschegg, S., Mayer, H., Wakita, M.
and Tanaka, K., Int. J. Fatigue, 30, 2057-2063, 2008.
7) Shimamura, Y., Narita, K., Ishii, H., Tohgo, K., Fujii, T.,
Yagasaki, T., Harada, M., Very High Cycle Fatigue
Properties of Carburized Steel by Ultrasonic Torsional
Fatigue Testing, 59, 938-943, 2010.
8) Pilkey, W.D., Peterson's Stress Concentration Factors(2nd Ed.), John Wiley & Sons, 128, 1997.
9) Ishii, H., Yamanaka, K. and Tohgo, K., Material
Science Research International, Special TechnicalPublication 1, 59-63, 2001.
10) ISO 281, Rolling Bearings−Dynamic Load Ratingsand Rating Life, 2nd Ed., 2007-02-15, Ref. No. ISO
281(E), 2007.
11) Shimizu, S., Tsuchiya, K. and Tosha, K., Tribol.
Trans. 52, 807-816, 2009.
12) Ouchida K., Journal of the Japan Society of
Mechanical Engineers, 64, 263-272, 1961.
13) Harris, T. A., Rolling Bearing Analysis, 3rd Ed., WileyInterscience, 147, 1991.

5. Conclusion
A typical example of stress among those governing
occurrence of initial shear-mode crack prior to
subsurface flaking failure on rolling bearings due torolling fatigue is repetitive orthogonal shear stress(near-purely alternating) that repeatedly acts on thesurface of the bearing. However, even if a rolling
bearing is tested under low loading conditions, asubsurface flaking failure will not occur during a limited
practical time frame. Though shear fatigue propertiesof rolling bearing steel can be evaluated throughtorsional fatigue testing, conventional fatigue testers
are only capable of applying low frequency loading to
specimens, and are virtually incapable of property
evaluation up to giga-cycles ranges.
To be able to quickly evaluate shear fatigue
properties of high-strength rolling bearing steels, NTN
has attempted to develop a unique ultrasonic torsional
fatigue tester (loading frequency 20,000 Hz,
alternating), which is capable of exerting shear stressamplitudes of up to approximately 950 MPa to
specimens. With this tester, NTN has successfully
obtained shear fatigue properties of high-carbon
chromium bearing steel JIS-SUJ2 from low cycle tohigh cycle regimes.

Photo of authors

Noriaki SAKANAKA

Yukio MATSUBARA

Yoshinobu SHIMAMURA

Hitoshi ISHII

Advanced Technology
R&D Center

Advanced Technology
R&D Center

Department of MechanicalEngineering,
Faculty of Engineering,
Shizuoka University

Emeritus Professor of
Shizuoka University-110-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Paper ]Two-Dimensional Dynamic Analysis of Cage Stress for Needle Roller
Bearings that Support Connecting Rods in Reciprocating Engines
Tomoya SAKAGUCHI*

In this paper, the author numerically analyzes cage stress in a connecting rod needle
roller bearing by using a dynamic rolling bearing model considering both roller and cage
motions as well as the cage elastic deformation. The analyses were conducted for four
conditions in a small motorcycle engine: maximum torque, maximum power, maximum
speed and overrun speed.
The results show that the cage stress rises at a moment immediately after ignition andduring the intake stroke when the crank angle is approximately 90 degrees from the top dead center position.
Additionally, it was determined that cage stress increases with the crank rotational speed.

forces and their resulting stresses within the needle
roller bearing. Fujiwara, et. al. built a test machine thatcan simulate the load acting on the connecting rod
large bearing and attempted to numerically determine
the behavior of the cage. This approach, however,does not cover orbital motions 1).
To cope with the difficulties in experimental
research into these motions, there have been
attempts to build dynamic models involving the rollers
and cage in order to determine numerically the timedependent motions of these components1)–3).
Regrettably, other than the areas in contact with therollers or raceway, the cage was regarded as a rigid
body, and stress acting on the cage was notconsidered in any of these models.
The author has attempted to research the motion ofrolling bearings through dynamic analysis 4) 5) while
being conscious about the cage's elastic deformation.
Consequently, the author has learned that in planetary
applications the stress occurring on the needle
bearing cage increases as the bearing speed
increases; an increase such that the calculated
maximum principle stress in the cage exceeds thefatigue limit of the cage material resulting in a cage
failure in a relatively short timeframe 5).
In this study, the author has applied a dynamic
analysis code 4), originally developed for planetary
applications, to analyze the effects of operating
conditions on crankshaft end, needle bearing cagestress.

1. Introduction
Single cylinder motorcycle engines adopt crankshaft
assemblies, and the interface between the crankshaft
and connecting rod (at the connecting rod large end)
often incorporates a needle roller bearing because its
starting torque is lower compared to a sliding bearing.
As demand for reduction in size and weight ofconnecting rods increases, the needle roller bearings
used for supporting connecting rods are required tohave greater load capacity. In order to increase load
capacity of a bearing within a limited space, increasein length and number of rollers is the primary solution.
However, this solution leads to thinner cage ribs
spacing the rollers, resulting in a decrease in the cage
mechanical strength. To address this issue, a cage
must be designed to have the mechanical strength
sufficient enough to avoid potential failure. Therefore,
it is necessary to calculate stresses on the cage withhigher accuracy.
The needle roller bearing motion at the crankshaft
end of the connecting rod is fairly complicated. The
inner ring raceway orbits around the center of
crankshaft while the outer ring raceway oscillates
relative to the inner ring raceway. The loads acting on
the bearing are pressure load from the piston, inertia
force from reciprocating motion and oscillating motionof the connecting rod, and centrifugal force from thecrankshaft rotary motion.
Because of such a complex combination of motions,
it is very difficult to measure in real-time the values of*Advanced Technology R&D Center-111-

NTN TECHNICAL REVIEW No.79（2011）

2. Analysis Method
The components of this analysis include the piston,connecting rod, crankshaft, and the mating needle
roller bearing as shown in Fig. 1. Major assumptionsof this analysis are as follows:
1. Three (3) degrees of freedom with rollers, cage andconnecting rod located on the yz plane.
2. Elastic deformation of the cage is considered using
the mode synthesis method i), 6).
3. The center axis of the crankshaft is fixed, and the
crankshaft turns at a constant speed.
4. The piston is linked to the connecting rod via asmoothly rotating joint and smoothly reciprocates in
the cylinder (friction is ignored while the piston isconstrained to move only in the y direction asshown in Fig. 1).
5. The rollers, cage, and raceway are mutually in anelastic contact state, and the normal force working
on them is defined by Hertz contact pressure.
6. Traction forces deriving from sliding on the raceway
surface and acting on the rollers in the rolling
direction are considered. Traction coefficient ismodeled using a function that reflects the effect ofproportion of oil film thickness to surface
roughness 7).
7. Rolling contact forces acting between the rollers
and raceway caused by the rolling viscosity dragoccurring from oil film 8)–10) and rolling direction
component occurring from pressure of oil film 8), 10)
are considered. These forces are determined by
the calculation techniques described by Zhou, et.
al. 10)On the basis of these assumptions, the author hasmodeled the rollers, cage, connecting rod, piston, andcrankshaft, as well as the interfering forces working
between these components. As a result, the motionsof these components, the deformation of the cage,
and the resulting stresses are calculated.

Fig. 1 Objects being analyzed, and coordinate system used

Table 1 Major specifications of the engine bearing, andlubrication conditions
Case for 4-cycle single-cylinder 125 cc engine
Crank radiusConnecting rod length
Connecting rod mass
Connecting rod center of gravity position*(*: height relative to center of big end)
Connecting rod moment of inertia
Piston mass

27.6 mm
102.7 mm0.0552 kg
35.9 mm
202 kg･mm2
0.0480 kg

Bearing specifications
Bore diameter
Outside diameter
Roller diameter
Roller length
Number of rollers

3. Objects Being Analyzed
Table 1 summarizes the specifications of theengine and bearing and the lubrication conditions tobe analyzed.
Table 2 summarizes the operating conditions of theengine being analyzed. These conditions are four
typical operating conditions seen in actual motorcycle
engines. Fig. 2 shows a plot of the time-dependent
variation in piston load that occurs on the piston due

26 mm
33 mm
3.495 mm10.8 mm16

Lubricating oil, kinematic viscosity
Representative temperature

100.5 mm2/s @ 40˚C
11.8 mm2/s @ 100˚C
135˚C

Table 2 Operating conditions of the enginei) Mode synthesis method is a technique to reproduce acomplicated deformation by addition and/or subtraction of
simple deformation modes.

-112-

Operating condition

Engine speed
min-1

Max. piston loadkN

Max. torque

5,000

16.3

Max. output

7,500

13.9

Max. speed

9,500

9.7

Overrun

10,500

7.6

Two-Dimensional Dynamic Analysis of Cage Stress for Needle Roller Bearings that Support Connecting Rods in Reciprocating Engines

Max torqueMax power

15000

12000

Max speed
Angular Velocity min-1

Piston load N

Overrun

10000

5000

8000

4000

0

-4000
0.02

0.03

0.04

0.05

0.06

0.07

Time s

0
-360

Outer
Inner
Cage

-270

-180

-90

0

90

180

270

a) Angular velocities of the outer ring,
inner ring, and cage

360

Crank angle degree

Fig. 2 Piston loads under various operating conditions

20000

Piston load
Bearing load
Force N

15000

to combustion pressure. The “0 degree” position of the
X axis represents a crank angle corresponding withthe top dead center position.

Explosion

10000

5000

4. Results of Dynamic Calculation

0
0.02

0.03

0.04

0.05

0.06

0.07

Time sFig. 3 shows the time-dependent change in angularvelocities of the major components, the piston load,
the bearing load, the slip ratio on cage, and the
maximum principle stress on the cage undermaximum torque conditions. The variation in angularvelocity of the outer ring shown in Fig. 3a)corresponds with the variation in oscillating motion of
the outer ring. Also, this chart makes it apparent thatthe angular velocity of the cage varies insynchronization with variation in its oscillating motion.
Fig. 3c) illustrates the slip ratio of the cage. The slip
ratio of the cage is defined as a percent decrease in
the effective cage rotational speed relative to the cage
rotational speed when there is no slip on the raceway.
When comparing slip ratio with the load on the
bearing, the slip ratio is greater in a low load region asshown in Fig. 3b). Fig. 3d) plots maximum principle
stresses on various points of the cage versus time (foroverview of the record of maximum stresses over theentire area of the cage, the records for multiple points
are indicated) . “Com” in this chart stands forcompression stroke, “Exp” explosion stroke, “Exh”
exhaust stroke, and “Int” intake stroke. The principle
stress soars up at the moment at which ignition occurs
in the cylinder (0.024 s and 0.048 s), then gradually
decreases.
Figs. 4 through 6 are time-dependent plots of cageloads and principle stresses under each of three
additional operating conditions. Figures 4, 5, and 6
have higher speeds respectively. As indicated in theb) Piston load and bearing load

Cage slip ratio %

40
30
20
10
0
-10
-20
0.02

0.03

0.04

0.05

0.06

0.07

0.06

0.07

Time sc) Slip ratio of the cage

Maximum stress MPa

800
600

Exp.

Exh.

Int. Com.

400
200
0
-200
0.02

0.03

0.04

0.05

Time sd) Maximum principle stress on the cage

Fig. 3 Results of numerical analysis in the" Maximum-torque" conditions-113-

NTN TECHNICAL REVIEW No.79（2011）

20000

20000

Piston load
Bearing load

Piston load
Bearing load

15000

Force N

Force N

15000

10000

5000

10000

5000

0
0.06

0.066

0.072

0.078

0.084

0
0.012

0.09

0.017

Time s

0.022

0.027

a) Piston load and bearing load
800

Exp. Exh.

600

Int.

Com.

Maximum stress MPa

Maximum stress MPa

Exp. Exh.

400
200
0
-200
0.06

0.066

0.072

0.078

0.084

Int. Com.

600
400
200
0
-200
0.021

0.09

Time s

0.017

0.022

0.027

0.032

b) Maximum principle stress of the cageb) Maximum principle stress of the cage

Fig. 4 Results of the numerical analysis in"Maximum-power" conditions

Fig. 6 Results of the numerical analysis in"Overrun" conditions

force versus time graphs, when the rotational speed
increases the bearing load during non-power strokes
will be greater. This force increase is because of theinertia force of the piston and connecting rod. The
direction of this inertia force at top or bottom deadcenter is away from the crank center: consequently,
the direction of inertia is opposite to the explosion.
Therefore, the bearing load at ignition is relatively
low while the engine is running at higher speed. Withan increase in rotational speed, the stress on the cageduring the intake stroke will be gradually increased.
Likewise, with an increase in rotational speed, thestress on the cage at ignition will be greater.
Nevertheless, the increase in stress at ignition is
generally small compared with that during the intake
stroke.

Piston load
Bearing load
15000

10000

5000

0
0.012

0.017

0.022

0.027

0.032

0.037

Time sa) Piston load and bearing load

800

Exp. Exh.

Int.

Com.

600
400
200
0
-200
0.021

0.017

0.022

0.027

0.037

Time s

20000

Force N

0.037a) Piston load and bearing load

800

Maximum stress MPa

0.032

Time s

0.032

0.037

Time sb) Maximum principle stress of the cage
Fig. 5 Results of the numerical analysis in"Maximum-speed" conditions-114-

Two-Dimensional Dynamic Analysis of Cage Stress for Needle Roller Bearings that Support Connecting Rods in Reciprocating Engines

5. Reasons for Increased Cage Stress
As a result of the analysis above, the author has
estimated that there are two major causes of cagestress increase:

Bearing load N

20000(1) The roller which has started orbital motion without
slip pushes the cage.
(2) The roller which has left the load zone, at aroundthe 90-degrees position in the intake stroke (crankangle relative to the top dead center) is acted
upon by the centrifugal force and hits a cage rib.

Com./Exp.
Ext./Int.

15000
10000
5000
0

Com./Exp.
Ext./Int.

Cage slip ratio %40

The reasons for increased cage stress will befurther discussed while referring further analysis
results.
Fig. 7 shows the bearing loads, cage slip ratios,and maximum principle cage stress versus the crankangle under maximum torque (lowest speed) andoverrun (highest speed) conditions . A crank angleof 0 in these plots corresponds with top dead center,“Com./Exp.” is the period from compression stroke topower stroke, and “ Exh./Int.” is the period fromexhaust stroke to intake stroke.
Under maximum torque conditions as shown in
Fig. 7a), the cage slip ratio increases immediately
before ignition, and suddenly drops immediately
afterwards. During this time, the maximum possible
stress occurs owing to reason (1) mentioned above.
The slip ratio of the cage will be highest at aroundthe 90 degree point in the intake stroke, and a spike
exists in the of cage stress plot. This peak stems
from reason (2) mentioned above.
Fig. 8 is a snapshot of the bearing motion at the
moment where the maximum principle stress occurs
during the maximum torque condition (crank angle
10.5 degrees). This figure illustrates the moment
immediately after the top portion of bearing hasentered the load zone due to combustion within theengine; thus the maximum principle load is occurringin the load zone. In this figure, the crankshaft andcage are rotating counterclockwise, the color contourof the cage indicates the magnitude of maximum
principle stresses, and the piston (not shown) is
situated above the diagram. The roller situated atthe 2 o’clock position is applying a counterclockwise
force to the corresponding cage rib. Fig. 9a)schematically illustrates this roller and its
surrounding cage ribs and shows the pre-ignition
situation. Due to inertia of the piston and connecting
rod, the rollers in the upper half of the diagram are
not subjected to load from the raceway; these rollers
are pushed away (in the upward direction in the
diagram) relative to the crank center owing to the

30
20
10
0
-10

Maximum stress on the cage
MPa

-20
800

Com./Exp.
Ext./Int.

600
400
200

0
-180

-90

0

90

180

Crank angle degreea) Maximum-torque conditions

Bearing load N

20000

Com./Exp.
Ext./Int.

15000
10000
5000
0

Com./Exp.
Ext./Int.

Cage slip ratio %

40
30
20
10
0
-10

Maximum stress on the cage
MPa

-20
800

Com./Exp.
Ext./Int.

600
400
200

0
-180

-90

0

90

180

Crank angle degreeb) Overrun conditions

Fig. 7 Bearing load, cage slip ratio and maximum
principle stress on the cage relative to the crank angle-115-

NTN TECHNICAL REVIEW No.79（2011）

centrifugal force ii) on the rollers generated by theorbital motion of the bearing. Consequently, the roller
at the 2 o’clock position comes into contact with thedownstream cage rib during rotation causing the rib todeform. Fig. 9b) indicates the situation at the moment
the explosion force has begun to act on the raceway.
Indicating the effect of the piston load, the upper
portion of Fig. 9b) shows the load zone. Being
subjected to the sudden increase in loading from the
raceway surface, the roller will begin to rotate rather
than slide: consequently, this roller pushes the cage rib
with a greater force, further deforming it. This
describes the increase in cage stress owing tomechanism (1) above. The most likely reason why theroller at the 2 o’clock position generates the maximum
stress is the simultaneous fulfillment of two conditions:first, the magnitude of the roller force (this force derives
from centrifugal force occurring on the roller itself)
acting on the cage rib is relatively large as shown in
Fig. 9a); and second, the magnitude of roller load fromthe explosion is relatively large, thus the frictional force
from the raceway contact is large.
Within the results of the overrun-speed conditions
shown in Fig. 7b), the maximum principle load on thecage occurs at around the 90-degree point during theintake stroke. This results from cage stress increasemechanism (2) described previously. Fig. 10 (intake
stroke, 89.2 degrees after top dead center) shows thebearing in this situation. In Fig. 10a), the roller at the 5o’clock position, situated outside the load zone,
collides with a cage rib. In this situation, the maximum
stress occurs in the corner of the cage pocket, asshown in Fig. 10b). This peak stress location coincides
with the usual failure point of standard cages.
Now, referring to Fig. 11, cage stress increasemechanism (2) is further described. Fig. 11 is aschematic rendering of the roller in the pocket where
the maximum stress has occurred as described andshown in Fig. 10. In Fig. 11 the piston is situated in theupper right, while the load zone is situated on the right.
Fig. 11a) shows a snapshot just prior to the occurrenceof the maximum principle stress. Consequently, theroller in the load zone at the 5 o’clock position applies
a force accelerating the rotation of the cage towards
the downstream cage rib. Fig. 11b) shows a snapshot
immediately after the depiction in Fig. 11a); wherein
the maximum principle stress has occurred. Since theload zone shifts at a velocity greater than velocity ofroller orbital motion, the roller is free from loading.

Fig. 8 Configuration of the bearing where the maximum
principle load acts on the cage under maximum torque conditions(explosion stroke, 10.5 degrees from top dead center)a) Roller and cage in the unloaded zone
immediately before explosionb) Force acting on the cage rib at the
moment of explosionii) The centrifugal force deriving from rotation of the roller aroundthe bearing acts vertically onto the outer ring raceway; thus it
does not affect rotation of the roller nor the magnitude of force
acting on the cage.

Fig. 9 Stress increase mechanism for the cageat the moment of explosion-116-Two-Dimensional Dynamic Analysis of Cage Stress for Needle Roller Bearings that Support Connecting Rods in Reciprocating Engines

Consequently, this roller is forced toward the left by
the centrifugal force due to orbital motion of the
bearing; thereby this roller collides with the upstream(rear) cage rib. This collision force results in an
increase of stress located at the cage pocket corner.

6. Effects of the Combustion Force andCrankshaft Rotational Speed on the
Cage Maximum Principle Stresses
The maximum principle stresses occurred on thecage under the various operating conditions shown in
Table 2. These stresses have been rearranged
relative to the rotational speed of the crankshaft. Fig.
12 graphically plots these results (“Each condition”).
Under different piston loads the stress increases with
the increase in rotational speed; however, there is no
linear relation between the rotational speed andstress. This appears to be because the maximum
torque condition (5,000 min-1) and the maximum
output condition (7,500 min-1) are relatively low speed
conditions, therefore maximum principle stress occurs
due to mechanism (1) above. Similarly, under the
maximum speed condition (9,500 min-1) and theoverrun condition (10,500 min-1), each being a highspeed condition, maximum principle stress occurs due
to mechanism (2) above. To verify this theory two
scenarios were reviewed, where the piston load was
fixed at zero (0 N, "None") and where the piston load
was set using the maximum torque condition(maximum load 16.3 kN, “Max. Torque”). Using theseloads, stresses under varying rotational speeds were
calculated, and the results have been added to theplot in Fig. 12. It should be understood from thisdiagram that when no piston load is applied (“None”),
the maximum principle stress on the cage increases inproportion to the 1.4th power of the crankshaft speed.

a) Front view of the overall bearingb) Area where the maximum principle load occurs

Fig. 10 Situation where the maximum principle stressoccurs on the cage under the overrun condition(air intake stroke, 89.2 degrees after top dead center)a) Roller and cage ribs at the edge of the load zone

Each condition
Max.TorqueNone
Maximum stress MPa

800

600

400

200

y=0.00164x1.404000b) Situation where the roller hits the outer cage rib
due to centrifugal force

6000

8000

10000

Rotation speed min-1

Fig. 11 Stress rise mechanism on the cage at around the
90 degree point after top dead center

Fig. 12 Maximum principle stresses on the cage
relative to the crankshaft rotational speed-117-

NTN TECHNICAL REVIEW No.79（2011）

This is because the stress occurring in accordancewith mechanism (2) is governed by the crankshaft
speed. On the other hand, when the piston load is set
under the maximum torque load (“Max. Torque”), thestress is greater over the entire speed range. The
bearing load resulting from the inertia of the pistonand connecting rod during the overrun (highest speed)
condition stands at 9 kN (bearing load value at top
dead center during exhaust or intake stroke in Fig.
6b), while the preset piston load (16.3 kN) is muchgreater. Due to this fact, the cage stress resulting frommechanism (1) increases beyond the maximum stressoccurring due to mechanism (2).

NTN remains committed to the analysis andverification of the maximum stress on the cage undervaried operating conditions through experiments;
resulting in cages designed to improve reliability andoptimize bearing performance.

Reference
1) Fujiwara, H., Kobayashi, T., Fujii, K., Dynamic Analysisof Radial Needle Roller Bearings, Proceedings of theTokai Branch Regular Meeting of the Japan Society of
Mechanical Engineers, 51, 83, 2002.
2) Walters, C.T., The Dynamics of Ball Bearings, ASME
J. Lub. Tech., 1, 1, 1971.
3) Gupta, P.K., Advanced Dynamics of Rolling Elements,
Springer-Verlag New York Inc., 1984.
4) Sakaguchi, T. et al., Dynamic Analysis of Cage Stressin Needle Roller Bearings under Planetary Motions,
Proc. ASME/STLE International Joint Tribology
Conference 2007, 44189, 2007.
5) Sakaguchi, T., Dynamic Analysis for Needle Roller
Bearings Under Planetary Motion, NTN Technical
Review No. 75, p. 94, 2007.
6) Craig, R.R., Bampton, M.C., Coupling of
Substructuring for Dynamic Analysis, AIAA 6, 7, 1313,
1968.
7) Sakaguchi, T., Harada, K., Dynamic Analysis Cage
Behavior in a Tapered Roller Bearing, ASME J. Trib.,
128, 604, 2006.
8) Houpert, L., Leenders, P., A Theoretical andExperimental Investigation into Rolling Bearing
Friction, Proc. 1985 Eurotrib Conf., 1, Session III,
Paper 2.8, 1985.
9) Aihara, S., A New Running Torque Formula for
Tapered Roller Bearings Under Axial Load, Trans.
ASME, J. Trib, 109, 7, 471, 1987.
10) Zhou, R.S., Hoeprich, M.R., Torque of Tapered Roller
Bearings, Trans. ASME, J. Trib, 113, 7, 590, 1991.

7. Conclusion
A two-dimensional dynamic analysis has been
completed for a needle bearing used at the crankshaft
end of a connecting rod, which considered motion of
the rollers, cage, and connecting rod as well as theelastic deformation of the cage, in order to study effectof operating conditions on the cage stress. The resultsof the analysis can be defined as follows:¡The following two mechanisms exist which increase
the cage principle stress:(1) Interference force between the rollers and cage
deriving from a rapid change in slip ratio that
results from the combustion force.
(2) Impact of a roller with the cage when the roller istransferred from the load zone to unloaded zone,
and then accelerates due to the centrifugal force.
¡Even when the crankshaft is turned without
involving the combustive force, the maximum
principle stress on the cage increases in proportion
to the 1.4th power of crankshaft rotational speed.
When combustive force causes piston load to occur,and if the piston load is greater than the sum of theinertia of the piston and connecting rod, then the
stress on the cage will further increase.

Photo of author

Tomoya SAKAGUCHI
Advanced Technology
R&D Center-118-

NTN TECHNICAL REVIEW No.79（2011）[ Technical Article ]Integrated Bearing Dynamic Analysis System (IBDAS)

Mariko SEKIYA*NTN has developed a new analysis system, Integrated Bearing Dynamic
Analysis System (IBDAS), in 2011. IBDAS is a dedicated platform for
dynamic rolling bearing analysis that is able to consider an elastic cagemodel and rolling elements with 6 degrees of freedom. Incorporating
automation and a user-friendly computer interface, a large reduction in
operation time can be achieved.
This paper compares static and dynamic analyses, provides a brief
explanation of IBDAS, and gives examples of completed dynamic analyses.

through repeated calculation often utilizing the
Newton-Raphson method. In bearing analysis,
centrifugal force and gyro moment due to rolling
element orbital motion are often considered; this issometimes referred to as quasi-static analysis.
Static analysis is very useful in estimating load
distribution, rigidity, and life of the bearing, providing
an indispensable means for rolling bearing design.
Conversely, static analyses are not capable ofcalculating frequently varying physical quantities in a
non-stationary state. In particular, it is difficult to obtainnecessary solutions for motion and stresses of abearing retainer with static analyses alone. These
factors cannot be easily measured experimentally;
there are several reports about static analysis forthese factors 1) 2) with limited measurement conditions.
Additionally, the scope of static analyses is limited
for bearings subjected to complicated external
environments, such as bearings used under varying
load and speed conditions. In cases such as these,
connecting rod bearings within an engine one
example bearing, studies need to be supported bydynamic analysis.

1. Introduction
In 2011, NTN began promoting its unique dynamic
analysis system for rolling bearings called IBDAS(Integrated Bearing Dynamic Analysis System) that isa dynamic analysis tool capable of not only
accommodating six degrees of freedom for the rolling
elements and cage (three transitional and threerotational), but also outputting the stress distribution
resulting from cage elastic deformation. Adoption of
automatic analysis and a refined computer interface
allows even new users to be able to execute a roller
bearing dynamic analysis.
This article provides an overview of the advantagesof dynamic analyses in comparison with static
analyses, describes the features of IBDAS, andprovides examples of completed dynamic analyses.

2. Features of Static and Dynamic
Analyses
Mechanical analysis for bearings can be categorized
into static analysis and dynamic analysis. Features
and applications of each analysis type are summarized
hereunder.

2.2 Dynamic analysis
Dynamic analysis is a technique where an equationof motion is developed for each element analyzed
within the mechanical system. Thus, simultaneous
ordinary differential equations are integrated overtime. Numerical results are considered using
numerical integration techniques, such as the RungeKutta method.

2.1 Static analysis
Static analysis is an analytical technique fordetermining various physical quantities in a stationary
state by solving a system of equations, typically bybalancing forces and moments. Generally these nonlinear simultaneous algebraic equations are solved*Automotive Business HQ. CAE Dept

-119-

NTN TECHNICAL REVIEW No.79（2011）

the contact pressure is calculated by Palmgren’s
simple formula 5).
･ In the case of a ball bearing, contact pressure
distribution is determined according to Hertz's
contact theory.
For calculation of the tangential force between
rolling element and raceway surface, the lubricant film
parameterΛis evaluated; thereby the relevant regimeof boundary lubrication, mixed lubrication andhydrodynamic lubrication are determined. Lubricant
film parameterΛis defined as h σ/ , which is the ratio of
the minimum lubricant film thickness (h) to meansquare surface roughness (σ) 6).
The tangential force coefficient μbd used in theboundary lubrication regime can be determined by
Kragelskii's 7) formula (1) as follows:

Compared with static analyses, computational time
for dynamic analyses is much greater. Nevertheless,dynamic analyses provide a useful tool for real-time
simulation, allowing one to analyze frequently varying
conditions, internal cage stresses, interaction forces
between bearing components, and physical quantities
such as rolling element skew and tilt.
One example of a rolling bearing dynamic analysis
technique is ADORE (Advanced Dynamics of Rolling
Elements) developed by Gupta 3); there are several
reports of calculation with this technique 4). However,
ADORE has drawbacks in that it can only handle a
cage as a rigid body rather than an elastic body, andthe scope of cage geometry ADORE handles islimited.
Table 1 summarizes features of static and dynamic
analyses.

μbd＝(-0.1+22.28s) exp(-181.46s)+0.1 …（1）s : slip rate

Table 1 Comparison of static and dynamic analyses

Staticanalysis

Constructive
formula/methodof solution
Force-momentbalance formula
Solution of algebraic
equation

Typicalcalculation
technique

Typicalapplications

Calculation
Time

NewtonRaphson
methodLoaddistribution
RigidityLife

Relatively
low

Non-static behavior
Motion equation
Dynamic (translational, rotational) Runge-Kutta In particular,
behavior andNumerical integration
methodanalysiswith differential equationstress of cage

The tangential force coefficient μbd used in thehydrodynamic lubrication regime can be determinedby a simple theoretical formula proposed by Kimura,et. al. 8). The mixed lubrication regime is expressed byformula (2) and provides an interface that smoothly
connects the boundary lubrication regime tohydrodynamic lubrication regime.

High

μr＝

｛

3. Features of IBDASIBDAS is a dynamic analysis system for rolling
bearings that enables automated calculations by using
commercially available dynamic analysis software.
This system features a unique tribological contact
model, an ability to handle the cage as an elastic
body, and an automated data processing userinterface. There features are briefly describedhereunder.

μbd-μhd(0.01-1.5)6μbd

if Λ＜0.01
6(Λ-1.5) +μhd if 0.01≤Λ＜1.5 （2）μhd

if 1.5≤Λ

As to the contact mode between the rolling element
and raceway surface, rolling viscous resistance 9)needs to be taken into account.
Tangential force factors for the forces acting
between the rolling elements and cage, as well as
between the raceway rings and the cage aredetermined by Formula (1).
For more detailed information about the tribology
model used for IBDAS, refer to the paper by Harada
and Sakaguichi 10).

3.1 Tribology modelContact modes within a bearing cannot be directly
expressed with a standard function available withincurrent commercially available, general-purpose,mechanics analysis software because of the presenceof oil film. Therefore, a specifically designed sub
routine was constructed.
Assuming that the contact mode between the rolling
element and raceway surface is localized elastic
contact, then, the contact pressure at the contact pointis calculated in the following manner:･ In the case of roller bearings, a slice method isapplied; thereby on the basis of each slice's load,

3.2 Elastic cage model
CMS method (Component Mode Synthesis) is used
to build an elastic cage model. This is a technique forexpressing the magnitude of deformation on an elastic
body by linear superposition of given deformation
modes 11).
Deformation modes are generated through analysis(known as super-element analysis) to obtain a minuteseries model by processing a characteristic (mass,attenuation and rigidity) matrix deriving from a finite-120-

Integrated Bearing Dynamic Analysis System (IBDAS)

element cage model on the basis of Guyan static
reduction 12). Then, on the basis of the deformation
mode information, stresses and deformation of thecage are calculated for the duration of the duty cycle
within the dynamic analysis software. A finite element
model which has not yet undergone static reduction
has degrees of freedom modeled with more than ten
thousands nodes. In contrast, NTN’s elastic bodymodel helps reduce the degrees of freedom down toscores of deformation models; thereby the work load
for the calculation is significantly reduced.
In static reduction, degrees of freedom for specified
nodes are left unchanged. Choice of these nodes(hereafter referred to as “boundary points”) will greatly
affect accuracy of the calculation: therefore, it ispreferable to choose points where external forces areapplied, where significant deformation is expected, or
where significant mass is present 12).
For correlation, NTN compared the behavior of atapered roller bearing retainer obtained by the mode
synthesis method with experimentally obtained
results 13).
In the experiment, a displacement sensor was
installed onto the test bearing (tapered roller bearing,
32310U), and the test bearing was run at 1000 min-1;
thereby the cage position was measured while varying
axial load. Fig. 1(a) shows the measurements
resulting from the experiment, and Fig. 1(b)graphically plots the results of calculation under thesame conditions. Axial load Fa is made dimensionlessby dividing by the basic dynamic load rating Cr, while
the cage position from the center is rendereddimensionless by ratioing to the radial clearance
between the cage pocket and roller.
Figs. 1(a) and 1(b) indicate that when the axial load
is low (Fa Cr=0.5%), the center of the cage shifts
forward, in the direction motion and away from thedirection of gravity. Additionally, when the axial load isgreater (Fa Cr=8%), the center of the cage traces anear perfect circle. Thus, the experimental results ofthe cage behavior qualitatively matches thecalculation results.
On the other hand, measurements of cage stressobtained from IBDAS are always lower compared withmeasurements obtained from the static analysis finite
element method (hereinafter referred to as FEM).
According to the dynamic analysis results under theconditions in Fig. 1, the cage stress can varydepending on the number of characteristic
deformation modes: nevertheless, the calculated cagestress values resulting from the dynamic analysis areapproximately 25% lower compared with stress values
obtained from the FEM 13).
To address this problem within IBDAS, a correction
coefficient for cage stress is calculated based on the
results of the static analyses performed by the FEM
and CMS methods.
Additionally, there has been concern about theunderestimation of stress concentration at the cage
pocket corner, wherein this underestimation can occur
when the roller-to-cage contact location deviates fromthe midpoint 14).
To address this problem, IBDAS is capable ofincreasing the number of boundary points on the cage
pocket rib. Fig. 2 shows the locations of boundary
points set on a cage pocket in a tapered roller bearing(equivalent to 30310). In the test results shown in Fig.
3, the number of boundary points set on the cage
pocket rib was varied, and then a radial load of 1 N
was applied to the loading points (points 1 through 6)
shown in Fig. 2. Fig. 3 graphically shows the ratio of
pocket corner stress obtained by CMS to cage cornerstress obtained by FEM. For more detailed information
about the analysis, refer to bibliography source 14.
From Fig. 3, it should be understood that when thenumber of boundary setting points is 1, and if load1

Perpendicular displacement

Perpendicular displacement

1

Fa/Cr=8%0

Fa/Cr=0.5%

Fa/Cr=8%0

Fa/Cr=0.5%

Rotation direction -1
-1

0

1

Rotation direction-1

Lateral displacement(a) Experimental result01

Lateral displacement(b) Calculation result

Fig. 1 Cage center behavior of a tapered roller bearing (32310U, 1000 min-1)-121-

NTN TECHNICAL REVIEW No.79（2011）acts on load points 1 or 6, each of which is in close
vicinity to the cage pocket corner, values resulting
from the CMS method are significantly lower
compared with values resulting from the FEM
analysis. With an increase in number of boundary
setting points, this trend is less apparent: if two or
more boundary points are set, the minimum “CMS to
FEM” stress ratio will be not smaller than 0.5.

4. Computer Interface
One outstanding advantage of IBDAS is the adoption
and ease of use of the computer interface. The computerinterface helps simplify input/output operations andallows for the automation of the calculation processing
section; thereby even new users of the software caneasily analyze the behavior of rolling elements.
Through the computer interface IBDAS operation
begins with the user entering a total of about 30 to 40
parameters including bearing type and size, lubricating
oil type, temperature, operating conditions such as load
and bearing speed, as well as numerical integration
conditions. Note, however, that the user needs todevelop a bearing mode shape file and upload the fileduring the data entry described above.
Once the user has completed data entry for all thecondition parameters, IBDAS then automaticallycalculates the cage deformation mode information,
constructs bearing model, and executes dynamic
calculation.
When the calculation operation is complete, the user
downloads the necessary data via the server andanalyzes the results. Macros are provided for theacquisition of frequently referenced data including the
behavior and stresses of the cage, contact force
between associated bearing parts, and roller skew andmisalignment; thus, the user can output necessary data
without experiencing complicated post-process
operation.
Fig. 4 illustrates an example of the work load
reduction through automation and utilization of the newinterface. Time-consuming data entry work and setup
conditions are no longer necessary. At the same time,
automated execution of calculations helps reduce thetotal work hours from 43 to a mere 2 hours. The work
load has been dramatically reduced to 1/20 of what it
was previously.
IBDAS can handle all rolling bearing types (deepgroove ball bearing, angular ball bearing, cylindricalroller bearing, needle roller bearing, tapered roller
bearing, and spherical roller bearing). IBDAS is also
capable of model selections based on operating
conditions—the user can select from a two-dimensionalmodel (planar constraints: two degrees of freedom fortranslational motion, one degree of freedom for
rotational motion) and a three-dimensional model (threedegrees of freedom for translational motion, threedegrees of freedom for rotational motion). Operation ofplanetary systems, eccentric rotation, and motion ofconnecting rods, etc. can be analyzed.
The user may also prepare a user-defined application
file, such that IBDAS can flexibly handle operating
conditions not listed above.

Large end
cornerSmall end
cornerFig. 2 Load points (1 through 6), boundary setting points(S, C, L), and stress measurement points(large end corner, and small end corner) on the cage

1.25

1P, Big end corner

1P, Small end corner

2P, Big end corner

2P, Small end corner

3P, Big end corner

3P, Small end corner3

5Stress ratio

1.00
0.75
0.50
0.25
0
1

2

4

6

Loading point

Fig. 3 Ratio of cage stress determined by CMS to cagestress determined by FEM
1P: One boundary point (C), 2P: Two boundary points (S, L),
Three boundary points (S, C, L)
Large end corner; Stress on the large end corner in cage pocket(Fig. 2, right)
Small end corner: Stress on the small end corner in cage
pocket (Fig. 2, left)-122-Integrated Bearing Dynamic Analysis System (IBDAS)

Necessary user’s time 43h
15h→1h2hInput

Data input

FEM analysis
Calculation

Integrated
automation tool

18h
→

User

0h
Dynamic analysis
Analysisof result

Analysis of calculation result

10h→1hFig. 4 Flow of dynamic analysis and example of time reduction due to IBDAS (light blue: previous system, pink: IBDAS)

method was incorporated into the two-dimensional
analysis operation for analyzing needle roller bearings
undergoing planetary motion 22). Eventually, NTN
attempted to verify the accuracy of the mode
synthesis method through the three-dimensional
analysis of a cage in a tapered roller bearing 13) 14).
In 2011, the dynamic analysis techniques forvarious rolling bearing types were combined andorganized into a unique user-friendly system—IBDAS.
For further example, the readers of this article areencouraged to refer to “Two-Dimensional Dynamic
Analysis of Cage Stress for Needle Roller Bearings
that Support Connecting Rods in Reciprocating
Engines” in this issue of the NTN Technical Review.

5. Disclosed Examples of NTN's
Dynamic Analysis
In order to present typical examples of theapplication of dynamic bearing analyses, the section
hereunder describes NTN’s history of dynamic
analyses up to the release of IBDAS. Note that theyear indicated is the year the information was firstpublished.
First, in 2000, using NTN’s proprietary developed
codes, the cage behavior and forces acting on thecage within a connecting rod needle roller bearing
were calculated 15) 16) and compared with theexperimentally determined cage behavior. This
analysis was focused on the large end connecting rod
needle roller bearing, which is a typical example ofdynamic operating conditions. This model was a twodimensional, simplified dynamic model whose degrees
of freedom was decreased with planar constraints.
Eventually, a commercially available dynamic
analysis software package was obtained. In 2002,
NTN disclosed results of its two-dimensional analysis
of cylindrical roller bearing cage behavior 17 18). NTN
also released results of its three-dimensional, rigidly
modeled cage analysis (an evolution from the aforementioned two-dimensional analysis) for a tapered
roller bearing 10) 19) 20); wherein NTN attempted toreview contact forces between the rollers and theretainer used in NTN’s newly developed bearing.
In 2006, more advanced three-dimensional
analyses became possible, which allowed the cage tobe handled as an elastic body using the mode
synthesis method 21). Eventually, the mode synthesis

Table 2 Progress in technique for dynamic analysis forrolling bearings at NTN
Year of outside
publication

Contents of publication

2000

Analysis of a needle roller bearing cage used in a connecting rod;according to NTN’’s propriety development code
(literatures 15 and 16)
Introduction of commercially available dynamic analysis software

2002

Two dimensional analysis of a cage (rigid model) in a cylindricalroller bearing using commercially available dynamic analysis
software (literatures 17 and 18)

2004

Three dimensional analysis of a cage (rigid model)in a tapered
roller bearing (literatures 19 and 20)

2006

Stress analysis of a cage (elastic model) in cylindrical roller
bearing, taking elasticity into account (literature 21)

2007

Two dimensional analysis of a needle roller bearing in planetary
motion (elastic model) (literature 22)

Release of integrated dynamic analysis system IBDAS
2011
2011年

-123-

NTN TECHNICAL REVIEW No.79（2011）14) Sakaguchi, T., Measurement Precision of Cage
Stress in Tapered Roller Bearings Using ComponentMode-Synthesis Method, Proceedings of Japan
Society of Tribologists Tribology Conference, 279-280,
2010.
15) Fujiwara, H. and Kobayashi, T., Dynamic Analysis of
Needle Roller Bearings in Connecting Rod Big-End
Applications, Proc. Int. Trib. Conf., Nagasaki 2000,
1841-1846, 2001.
16) Fujiwara, H., Kobayashi, T., Dynamic Analysis of
Needle Roller Bearings in Connecting Rod Big-End
Applications, NTN Technical Review No. 69, 89-96,
2001.
17) Sakaguchi, T., Ueno, K., Kobayashi, T., Analysis ofcage behavior on cylindrical roller bearings,
Proceedings of Japan Society of Tribologists Tribology
Conference, 415, 2002.
18) Sakaguchi, T., Ueno, K., Dynamic Analysis of Cage
Behavior in a Cylindrical Roller Bearing, NTN
Technical Review No. 71, 8-17, 2003.
19) Harada, K. and Sakaguchi, T., 3D dynamic analysis
of cage behavior on tapered roller bearings (1st report,
behavior measurements), Proceedings of Japan
Society of Tribologists Tribology Conference, 501,
2004.
20) Sakaguchi, T. and Harada, K., 3D dynamic analysis
of cage behavior on tapered roller bearings (2ndreport, calculation results), Proceedings of Japan
Society of Tribologists Tribology Conference, 503,
2004.
21) Sakaguchi, T. and Harada, K., Dynamic Analysis ofCage Stress in Tapered Roller Bearings, Proc.
ASIATRIB., Kanazawa 2006, 649-650, 2006.
22) Sakaguchi, T., Dynamic Analysis for Needle Roller
Bearings Under Planetary Motion, NTN Technical
Review No. 75, 94-99, 2007.

6. Conclusion
This article has compared the techniques of staticand dynamic analyses, explained the advantages ofthe dynamic analysis system IBDAS, and summarized
NTN’s typical experiences using dynamic analysis.
Through the adoption of dynamic analysis models,stresses and behavior of the bearing cage undervaried operating conditions are clarified (thisachievement was impossible using static analysis
techniques). At the same time, automation of theanalysis operation has helped to greatly decrease theanalysis time required.
NTN will promote utilization of both static anddynamic roller bearing analysis techniques in order tofurther enhance its ability to develop better bearing
designs and support its products.

Reference
1) For example: Kakuta, K., Kohno, T., Ohta, H., Ota, H.,
Generating Mechanism of Forces Acting on Retainer ofBall Bearings Supporting Unbalanced Rotating Shaft,Transactions of the Japan Society of MechanicalEngineers, 62-600, C 3210-3215, 1996.
2) Fujiwara, H., Fujii, K., Nakazeki, T., Retainer Forces in
Cylindrical Roller Bearings, NTN Technical Review No.
68, p. 63ﾐ66, 2000.
3) Gupta, P.K., Advanced Dynamics of Rolling Elements,
Springer-Verlag, 1984.
4) Gupta, P.K., Dynamics of rolling element bearings,
Parts I, II, III and IV, Journal of Lubrication Technology,
ASME Trans., 101, 293-326, 1979.
5) Gupta, P.K., Advanced Dynamics of Rolling Elements,
Springer-Verlag, 39, 1984.
6) Yamamoto, Y. and Kaneta, M., Tribology,
Rikogakusha, 162, 1998.
7) Kragelskii, I.V., Friction and Wear, Butterworths, 1965.
8) Muraki, M., Kimura, Y., Traction Characteristics of
Lubricating Oils (Part 2), Journal of Japan Society ofLubrication Engineers, 28, 10, 753-760, 1983.
9) Zhou, R.S. and Hoeprich, M.R., Torque of Tapered
Roller Bearings, ASME J. Tribology, 113, 7, 590-597,
1991.
10) Harada K., Sakaguchi, T., Dynamic Analysis of aHigh-Load Capacity Tapered Roller Bearing, NTN
Technical Review No. 73, 20-29, 2005.
11) Craig, R.R., and Bampton, M.C.C., Coupling
substructures for dynamic analysis, AIAA J., 6, 7,
1313-1319, 1986.
12) Nagamatsu, A., Okuma, M., Component mode
synthesis, Baifukan, 1991.
13) Sakaguchi, T. and Harada, K., Dynamic Analysis ofCage Stress in Tapered Roller Bearings Using
Component-Mode-Synthesis Method, ASME J.
Tribology, 131, 011102 1-9, 2009.

Photo of author

Mariko SEKIYA
Automotive Business HQ.
CAE Dept-124-

NTN TECHNICAL REVIEW No.79（2011）[ New Product ]Hybrid PEEK Sliding Bearing

Takuya ISHII*Ken YASUDA**

PTFE-based bushings are used for the main spindle bearings of air
conditioner compressors. Recently, demand has increased forcompressors with enhanced efficiency, reduced size, and lighter weight. To
meet these requirements, NTN has developed Hybrid PEEK Sliding
Bearings, which have injection molding PEEK resin layers of 0.5 mm
thickness on the bores of sintered metal bushings, with the objective ofimproving friction/wear properties and anti-seizure performance. This
article introduces the characteristics, structure and performances of Hybrid
PEEK Sliding Bearings.

1. Introduction
Improved energy efficiency is needed for bothhousehold and commercial air conditioners. Most of
the power consumption of an air conditioner is due tooperation of the air compressor. Therefore, the aircompressor in an air conditioner system needs to bevery energy efficient 1). The structure of a typical air
compressor for an air conditioner is shown in Fig. 1
along with the locations of bearings used. The main
spindle bearings (PTFE resin bushings) have tosatisfy requirements for improved friction-abrasion
quality and seizure resistance, because refrigerantliquefied by over-compression can cause refrigeration
oil to flow out via the friction surface. This article
presents information about NTN’s unique hybrid PEEK
bearing that hybridizes newly developed sinteredmetal alloy and resin material in order to meet theserequirements.

Fixed scroll
Swing scroll
Thrust bearing
Main spindle
bearing
Electric motor

Main spindle

Fig. 1 Typical structure of air-conditioner compressor

performance solves this issue with hybridization of athin-walled resin material and a metal layer. Thissliding bearing is a PTFE resin bushing 2) (Fig. 2) thatis used as the main spindle bearing in an air
conditioner compressor. However, because of alimitation in the manufacturing process, the resinmaterial used is limited only to PTFE. If refrigeration
oil flows out via the friction surface as in the case of
the compressor main spindle, the oil attacks the thin
resin layer , exposing the sintered metal layer and2. Performance Improvement for Resin
Sliding Bearing
Compared with rolling bearings, resin sliding
bearings are inferior in terms of heat resistance. Evena super engineering plastic material featuring
relatively high heat resistance can melt and wear out
due to poor heat releasing performance.
A resin sliding bearing boasting high heat releasing**Composite Material Products Division Composite Material Engineering Dept**NTN Engineering Plastics Corp. Engineering Dept-125-

potentially causing bearing seizure. Furthermore,
there have been mounting needs for lower bearing
torque and improved abrasion resistance. To address
this trend, NTN has developed the new hybrid PEEK
bearing.

PTFE resin

Sheet steel

Sintered
layer

0.5mm

NTN TECHNICAL REVIEW No.79（2011）

Sintered metal alloy bushing

PEEK resin

Resin
layer

Fig. 3 Structure of hybrid PEEK sliding bearing

Sheet steel

3.3 General physical properties

Cross-section

Table 1 summarizes general physical properties of
PEEK resin (BEAREE PK5307).

Fig. 2 Three-layered PTFE resin bushing(sheet steel + porous sintered bronze material + PTFE)

Table 1 General physical properties of BEAREE PK5307
Characteristics

Test method

Specific gravity

ASTM D792―

3.1 Features

Tensile strength

ASTM D638

MPa

83

Note that a conventional PTFE resin bushing isfabricated by impregnating a sintered intermediate
layer on sheet steel with PTFE resin and baking theintermediate layer. In contrast, with NTN’s new hybrid
PEEK bearing, a layer of 0.5 mm thick PEEK resin isformed onto the bore surface of a sintered alloybushing through an injection molding process. PEEK
resin excels in heat resistance, abrasion resistance,
oil/chemical resistance and fatigue resistance.
Hybridization with a metal layer helps achieve
dramatically improved heat releasing performancecompared with bearings consisting of PEEK resinmaterial alone.

Bending strength
Young’s modulus
in flexure
Coefficient oflinear expansion(R.T.〜150˚C)
Izod impact
strength

ASTM D790

MPa

142

ASTM D790

MPa

5,200

TMA method

1/˚C

MD: 3.4×10-5
CD: 5.3×10-5

ASTM D256(notched)

J/m

30

ASTM D785

R scale117

3. Hybrid PEEK Bearing

Rockwellhardness

Unit of measurement Characteristic value
1.42※All the values above are representative.

3.4 Comparison of various features of bearings
Table 2 compares NTN’s new hybrid PEEK bearing
product to a conventional PTFE resin bushing.
Comparison test data is given in Sec. 3.5, showing
NTN’s newly developed hybrid PEEK bearing excels
in seizure resistance and friction-abrasion
performance.

(1) The new product excels in seizure resistance(comparison with conventional product: five times or
more)(2) Lower friction/reduced abrasion quality(comparison with conventional product: 2/3 in termsof coefficient of friction, 1/3 in terms of abrasion)* Conventional product: PTFE resin bushing

Table 2 Comparison of various bearing features
Characteristics

Hybrid PEEK
bearing

PTFE resin-based
bushing

Surface layer(sliding surface)

PEEK resin(0.5mm)PTFE resin(0.05mm)Intermediate layer

None

Sintered bronze(0.3mm)Backing metal

Sintered metal alloy

Sheet steel

Form

Bushing(without split)

Bushing(with split)

Resin material forming
process

Injection molding

Impregnation forming

Seizure resistance◎△Abrasion resistance◎○Composition

3.2 Structure and constituent materials
Fig. 3 shows a simplified illustration of NTN’s newly
developed hybrid PEEK bearing. The PEEK resinmaterial (NTN material name: BEAREE PK5307)
forms the bearing bore surface that is slightly incontact with the matching shaft, and includes a blendof special filling material that boasts good frictionabrasion quality while still being capable of forming athin film through an injection molding process. The
bushing is made of a ferrous sintered alloy.

Friction resistance◎◎：Excellent

-126-

○
○：Good△：AcceptableHybrid PEEK Sliding Bearing

3.5 Bearing performance
3.5.1 Seizure resistance(4) A bearing temperature rise in excess of 10˚Cmarks the occurrence of seizure.

Seizure resistance testing was performed with thehybrid PEEK bearing, a PTFE resin bushing and a
PEEK resin bearing. Fig. 4 shows a schematic of theradial tester used, and Table 3 summarizes theapplied test conditions. The test procedure issummarized as follows:

Fig. 5 provides data about the time elapsed before
seizure with the tested bearings. NTN’s hybrid PEEK
bearing is superior to the PTFE resin bushing in termsof seizure resistance quality. As illustrated in Fig. 5,
the PEEK resin bearing, not hybridized with a metal
layer, seizes at an earlier stage of the seizure
resistance test.

<Test procedure>(1) Oil flows throughout the test rig and the shaft
rotates.
(2) Increase the bearing contact pressure by 0.8 MPa
every 3 minutes, so that the contact pressure
reaches 17 MPa, 1 hour after start of the test.
(3) When contact pressure of 17 MPa is reached, oil
circulation is stopped; oil is drained away from thebottom of the lubrication system, and then the
bearing runs until seizure (30 min. at maximum).

Load

3.5.2 Friction-abrasion characteristics
A radial friction-abrasion test was performed withthe hybrid PEEK bearing and a PTFE resin bushing.
These parts were subjected to a radial frictionabrasion test in lubricating oil under the test conditions
summarized in Table 4.
Fig. 6 shows the relationship between test duration
and wear depth, and Table 5 gives data for thecoefficient of dynamic friction. After 30 hours from the

Bearing

Oil circulation
port

Oil circulation
port
IN

OUT

Table 4 Test conditions

Matching shaft
Oil drain
port

Characteristics

Contents

Test machine

Shaft rotation radial test machine (Fig. 4)

Bearing size

Bore dia.: 30 mm, width: 20 mm

Operating internal clearance 0.09mm

OUT

Fig. 4 Radial-type test machine
Table 3 Test conditions
Characteristics

Contents

Bearing size
Operatinginternal clearance
Contact pressure

Bore dia.: 30 mm, width: 20 mm

Matching shaft

4.7m/s（3,000min-1）

Lubricating oil(circulation)

Ether-based refrigerating oill

Matching shaft

FCD600（0.4μmRa）(temperature: 90˚C, flow rate: 100 mL/min)

120h

50

max. 17 MPa (load 10 kN)-1）PTFE resin bushing

8.5m/s（5,400min
14.1m/s（9,000min-1）

40

Ether-based refrigerating oil(temperature at circulation: 90˚C, flow rate: 100 mL/min)

FCD600（0.4 μmRa）

Time elapsed after stopping of oil circulation,
before occurrence of seizureMin.
0

Sliding velocity

0.09mm10

20

Hybrid PEEK
bearing
PTFE resin
bushing

30

Longer than 30 minutes(Can continue operation.)

Lubricating oil

17 MPa (load 10 kN)

Duration

Wear Depth μm

Sliding velocity(two conditions)Contact pressure

Hybrid PEEK bearing
30

20
10
0
0

20

40

60

80

100

120

Time hFig. 6 Wear depth versus testing time
Table 5 Coefficient of dynamic friction

PEEK resin
bearing

Sliding velocity
8.5m/s14.1m/sFig. 5 Seizure resistance-127-

Bearing

Coefficient of dynamic friction(120 hours after start of test)

Hybrid PEEK bearing

0.006

PTFE resin bushing

0.009

NTN TECHNICAL REVIEW No.79（2011）

3.6 Environment resistance3.6.1 Heat cycles

start of the test, wear depth remains virtuallyunchanged. The hybrid PEEK bearing excels infriction-abrasion quality. Compared with the PTFE
resin bushing, the hybrid PEEK bearing boasts 2/3 thecoefficient of friction and 1/3 the wear depth.

To determine the variation in adhesion of PEEK
resin to the sintered metal alloy bushing resulting from
repetition of a thermal expansion-contraction cycle, ashear fracture test was executed with samples before
and after undergoing heat cycles (Table 7). As
illustrated in Fig. 8, shear force was applied to the
PEEK resin layer on the bearing bore surface. Then,the shear fracture load was applied to the area of thebore surface within the sintered metal alloy bushing inorder to determine the shear strength of the PEEK
bearing. Table 8 summarizes the shear stress values
obtained from NTN’s newly developed hybrid PEEK
bearing both before and after application of heat
cycles.
The results of the test determined that the PEEK
resin member did not change in external appearance,such as loosening from the sintered metal alloybushing. At the same time, virtually no deterioration
was found in shear strength. This is because of
improved adhesive power owing to the anchoring
effect of recesses on the surface of the sintered metalalloy bushing. The PEEK sliding bearing that
underwent the shear fracture test had a mixture offractured PEEK resin in recesses of the sinteredbushing and inter-surface fracture between the PEEK
resin and sintered metal alloy bushing.
Furthermore, the change in frictional shear strength

3.5.3 Effect of contact pressure on coefficientof friction
The effect of contact pressure on the coefficient offriction was tested for the hybrid PEEK bearing and a
PTFE resin bushing under the test conditions
summarized in Table 6. The contact pressure on thetest samples in lubricating oil was increased by 1 MPa
every minute, beginning with 1 MPa up to a maximum
pressure of 10 MPa. Fig. 7 graphically plots theresults of the effect of contact pressure on thecoefficient of dynamic friction.
Compared with a conventional PTFE resin bushing,
NTN’s newly developed hybrid PEEK bearing boasts
greater load capacity. Its coefficient of friction is about
2/3 that of the PTFE resin bushing. The coefficient offriction with NTN’s hybrid PEEK bearing tends to belower, along with lower contact pressure.

Table 6 Test conditions
Characteristics

Contents
Shaft rotation radial test machine (Fig. 4)

Test machine
Bearing size

Bore dia.: 30 mm, width: 20 mm

Operating internal clearance

0.09mmContact pressure

max. 1–10 MPa (load 6 kN)

Sliding velocity

4.7m/s（3,000min-1）

Lubricating oil(circulation)

Ether-based refrigerating oil

Table 7 Heat cycle test conditions(temperature: 90˚C, flow rate: 100 mL/min)

Matching shaft

FCD600（0.4μmRa）

Characteristics

Conditions

Pattern-40˚C (30 min.) ⇔ 165˚C (30 min.)Number of repetitions

50 times

Shear direction
Coefficient of dynamic friction

0.12

Jig

PTFE resin bushing

0.10

Sintered metalalloy bushingHybrid PEEK bearing
0.08
0.06

PEEK resin(thickness 0.5 mm)

0.04
0.02
0.00

0

2

4

6

8

10

Fig. 8 Schematic view of shear test rig

12Contact pressure MPa

Table 8 Result of shear test

Fig. 7 Effect of contact pressure on coefficient of friction

Before heat cycle(original state)

After heat cycle

Shear strength

3.62MPa

3.53MPa

Rate of change(Reference)-2.6%

Characteristics-128-

Hybrid PEEK Sliding Bearing

before and after heat cycles is approximately 0.1
MPa, which was calculated along with the results
shown in Fig. 7 (the product of contact pressure andcoefficient of friction). Thus, there are no problems
expected using the PEEK resin bearing in air
compressors.

oil immersion on shear strength and the bore
dimension of the PEEK resin bearing.

4. Conclusion
Compared with a conventional PTFE resin bushing,
NTN’s hybrid PEEK bearing boasts much improved
seizure resistance quality and friction-abrasion
characteristics.
NTN will attempt to further expand the scope of
possible applications of the hybrid PEEK bearing not
only to household and commercial air conditioner
compressors, but also to automotive air conditioner
compressors. At the same time, aiming at further
improvement in functionality and reduction in size andweight, NTN will mobilize all of its sliding bearing
technologies (resin bearings, sintered oil-impregnatedbearings, fluid dynamic pressure bearings) to develop
and commercialize highly functional bearing products,
positively contributing to energy savings, efficiency
improvement, and reduction in size and weight ofindustrial machinery.

3.6.2 Oil resistanceTo determine the change resulting from long use in
oil regarding the adhesion between the PEEK resinand sintered metal alloy, as well as the dimensionalvariation of the PEEK resin layer, NTN implemented
an oil immersion test under conditions summarized inTable 9. Figs. 9 and 10 summarize the relationshipbetween oil immersion time and shear strength, aswell as change in bearing bore diameter. No change
in appearance was found on the bearing samples that
underwent oil immersion. There is virtually no effect ofTable 9 Oil immersion test conditions
Characteristics

Conditions

Immersion oil

Ether-based refrigerating oil

Temperature

150˚C

Duration

100h，300h，500h

Reference
1) Sato, H. et al., High Efficiency and Large Capacity 3D
Scroll Compressor GU Series, Mitsubishi Heavy
Industries Technical Review, 43, 2, 10, 2006.
2) MLE bearing, Cat. No. 5116-II/J, NTN, 2006.

Shear strength MPa

5
4
3
2
1
0

0

100

200

300

400

500

Time hChange in bore diameter mm

Fig. 9 Relation between oil immersion time andshear strength

0.01

Photo of authors

0
-0.01
-0.02
-0.03
-0.04

0

100

200

300

400

500

Time hFig. 10 Relation between oil immersion time and change
in bearing bore diameter-129-

Takuya ISHIIKen YASUDA

Composite Material
Products Division
Composite Material
Engineering DeptNTN Engineering Plastics Corp.
Engineering DeptNTN TECHNICAL REVIEW No.79（2011）[ New Product ]Air Spindle for Ultra High-precision Machine Tools

Teruyoshi HORIUCHI*
Kazuyuki AONO*

Air spindles are used in applications such as inspection equipment for
semiconductor wafers and hard disks. At NTN, we are using our skills andexperience to produce air spindles for ultra-high-precision machine tools.

aberration, and provides a smaller and simpler
solution. However, unlike production of simple
spherical lenses which can be readily finished withlapping, small lot production of nonspherical lenses
involves first trace-cutting of multiple-curved faces witha cutting tool, and then finishing with lapping. The
usual mass-production technique for nonsphericallenses is forming with metal dies. For high-efficiency
forming, the metal dies must have accurately finishedinternal profiles consisting of complicated curved
faces. This requirement can be met with a machining
spindle which features minimized rotational runout anddimensional variation.

1. Introduction
Thanks to excellent rotary characteristics resulting
from non-contact construction, air spindles have been
contributing to performance improvement in various
industrial fields, including disk-type data storage
media production and semiconductor device
inspection. Furthermore, air spindles have been
increasingly regarded as a promising tool for accuracy
improvement in precision machining processes fornonspherical lenses and special metal dies.
This trend is promoted by users’ mounting demands
for digital cameras with higher functionality, cellular
phones with built-in ultra-compact cameras, andsmartphones, each of which require compact, highprecision nonspherical image-capturing lenses.
This paper describes typical performance &assessment results and an explanation of the control
scheme for NTN air spindles for machine tools,
targeting the production process for nonsphericallenses.

2. Nonspherical Lens
Generally, when configuring an optical system byusing spherical lenses of the single radius form, it isnecessary to combine multiple lenses to correctaberrations associated with individual lenses;
consequently, the construction of the optical system iscomplicated. In contrast, a nonspherical lens such asa one in Fig. 1 is formed with a plurality of sphericalfaces; therefore, even one spherical lens can minimize

Fig. 1 Nonspherical lens

* Aberration is a phenomenon where light having passed a lensdoes not converge on a narrow point on the optical axis, andthe resultant image appears to be blurred or distorted.

*Precison Equipment Division Product Engineering Dept

-130-

Air Spindle for Ultra High-precision Machine Tools

air spindle operation; (b) component design intended
to minimize heat buildup during operation; and (c)careful selection of materials for limited thermal
expansion properties.

3. Features of Air Spindles for Machine
Tools
Previously, air spindles were very often used in an
upright attitude. In contrast, NTN’s newly developed
air spindle is intended for horizontal installation suchas shown in Fig. 2. The major targeted applications of
this air spindle product are precision turning centersand grinding machines.
For evaluation purposes, we have fabricated a
prototype air spindle for a machine tool (hereinafterreferred to as “newly developed product”), and Table 1
summarizes the major specifications for this newly
developed product. The drive motor used is capable ofcontinuous operation at 20,000 min-1: the newly
developed air spindle boasts high degree of accuracy,
and features radial and axial NRRO (Non Repetitive
Runout) values of 0.010 µm or smaller. This air
spindle is more reliable, thanks to its non-contact typeseal that more reliably prevents possible ingress ofchips and coolant into the bearing. Furthermore, thisNTN product is less affected by variations intemperature. This is achieved thanks to several
innovations: (a) the well-thought-out cooling structure,
which helps mitigate heat buildup during continuous

4. Evaluation of Rotational Accuracy
NTN's typical air spindle evaluation method is
described below, along with actual accuracy
measurements of NTN's newly developed air spindle
product.

4.1 Non Repetitive Runout (NRRO) accuracy(1) Method for measurement and evaluation
The rotary shaft is allowed to run at a constant
speed. The runout of an arbitrary point on the rotary
shaft is measured when the rotary shaft has justcompleted one full rotation, and the difference
between this measurement and the measurement at
the start of the rotation is determined. Then, thisdifference is acquired for 2,000 successive rotations;the maximum value of this difference is taken as the
non-repetitive runout (NRRO) for evaluation purposesof the spindle accuracy.
The analyzer used is an electrostatic, capacitance

Variation on surface ofworkpiece being machined

Chuck
Target of spheremeasurement
Axial rigidity
measurement point
Axial NRRO
measurement point
Radial NRRO
measurement point
Radial rigidity
measurement point

Fig. 2 Accuracy measurement points
Table 1 Specifications
Characteristics(at bearing feed air pressure 0.49 MPa)

Guaranteed value

Max. rotational speed

20,000min-1

NRRO

Radial rigidity

45N/μmAxial rigidity

227N/μm

Radial

0.010 μm max.

Axial

0.010 μm max.

Rotational speed fluctuation (jitter)within ±0.01%

Variation on surface ofworkpiece being machined(at constant rotational speed)

1 μm max.

Fig. 3 Test rig for NRRO measurement-131-

NTN TECHNICAL REVIEW No.79（2011）(3) Drive system

type non-contact displacement gage (MicroSense
from ADE) with a measuring resolution of 0.001 µm,
and the accuracy was measured on a sphericalmeasurement target shown in Fig. 3.

To be able to run an AC motor, sinusoidal current
needs to be fed into its motor coil. The drive system
for an AC motor can be categorized into “linear drive
system” and “PWM (pulse width modulation) drive
system” types, depending on the method forgenerating the sinusoidal current.
With the linear drive system represented in Fig. 6, asinusoidal command voltage is power-amplified todrive the motor. Though boasting higher control
accuracy, this system involves greater loss at theamplifier; consequently, it is difficult to achieve larger
capacity—a drawback that needs to be solved.
With the PWM drive system represented in Fig. 7,
the pulse width of command voltage is adjusted in(2) Result of measurement
At the rotational speed of 20,000 min-1, radial andaxial NRRO were measured. Figs. 4 and 5 show theactual measured values. The new product,
incorporating the improvements described earlier—heat buildup prevention arrangement, sealing
performance of improved mechanisms including
improved chuck—has stably achieved runout of 0.007µm or less. Therefore, it has been determined that the
new product satisfies the guaranteed values.

Rotational speed: 20,000 min-1

0.020

0.018

Displacement μm

0.018

Displacement μm

Rotational speed: 20,000 min-1

0.020

0.016
0.014
0.012
0.010
0.008
0.006

0.016
0.014
0.012
0.010
0.008
0.006

0.004

0.004

0.002

0.002
0

0
0

200

400

600

800

1000

1200

1400

1600

1800

0

2000

200

400

600

800

1000

1200

1400

1600

1800

Number of measurements

Number of measurements

Fig. 4 Radial NRRO on linear drive system

Fig. 5 Axial NRRO on linear drive system

Commandvoltage

CommandvoltagePoweramplificationPoweramplification

Motor drive
current

Motor drive
current

Fig. 6 Linear drive system

Fig. 7 PWM drive system

-132-

2000

Air Spindle for Ultra High-precision Machine Tools(2) Evaluation results of jitter by drive system

proportion to the amplitude of the sinusoidal wave byswitching ON/OFF power, and then the voltage pulses
are power-amplified to generate the current thatapproximates the sinusoidal wave in order to drive themotor. This system is highly versatile and highly
efficient, and can readily offer higher output. However,noise occurring from switching can adversely affectaccuracy of the air spindle; consequently, this drive
system is inferior to the linear-type drive system interms of accuracy.
NTN's newly developed air spindle is combined witha dedicated high-capacity linear drive-type driver toachieve higher accuracy.

Magnitude of jitter was measured for both the linear
drive system and the PWM drive system. The
resultant actual measurements are summarized inTables 2 and 3. Jitter is smaller with the linear drive
system, and accuracy of PWM is inferior owing to theadverse effect of noise.

5. Functions and Construction Required on
Air Spindle for Machine Tool
(1) Heat buildup prevention features
Mitigation of heat buildup on air spindle is a difficult
challenge if the air spindle is used on a machine tool
intended for highly accurate machining operation. Thisis because the components in bearings within the air
spindle can expand or shrink due to fluctuation intemperature, and the resultant displacement will betransferred to the workpiece being machined. The
typical contributing factors for temperature fluctuation
can include fluctuation in room temperature, and heatgeneration from the drive and bearing. To mitigate
these heat buildup factors, the NTN’s newly developed
air spindle has adopted the following features:(4) Evaluation results of NRRO by drive system
NTN prepared two systems for comparison of
NRRO performance. Both used the samecombination of air spindle proper and electric motor,
while two different drive systems -- linear and PWM -were applied, and the NRRO compared between the
two. Fig. 8 shows plotting of measured radial NRRO
values obtained from the PWM drive system. With the
PWM drive system, the peak reached 0.018 µm,
owing to effect of switching noise; this means theamplitude of displacement with the PWM drive system
is more than twice as great compared with the linear
drive system in Fig. 4.

[1] Water cooling structure
A water jacket is adopted to enclose the entire
bearing system and motor for efficient cooling of the
air spindle.

Rotational speed: 20,000 min-1

0.020[2] Use of low thermal expansion materials

Displacement μm

0.018

Low thermal expansion materials such as invar areused for certain bearing components to mitigate
thermal displacement.

0.016
0.014
0.012
0.010
0.008
0.006

Table 2 Measurement result of jitter on linear drive system

0.004
0.002

JitterRotational speed
Displacement Δt
±%
min-1
nsec
213.0
180.1
190.9
205.5
195.0
0.00009
500

0
0

200

400

600

800

1000

1200

1400

1600

1800

2000

Number of measurements

Fig. 8 Radial NRRO on PWM drive system

4.2 Rotational speed fluctuation (jitter)(1) Method for measurement and evaluation
Jitter is a numerical measure for defining theseverity of rotational speed fluctuation. A control rotary
encoder is included in the air spindle. The origin pulse
from the rotary encoder (one pulse per rotation) is fed
into the analyzer used. Proportion of fluctuation inrotational cycle (Δt) to theoretical value (T) is referred
to as jitter and is used as a measure for determining
quality of rotational accuracy with the air spindle.

1800

82.1

72.7

68.5

84.3

67.6 0.00013

5400

33.5

28.2

27.9

31.8

26.5 0.00015

10000

20.8

21.2

17.4

21.2

18.8 0.00018

15000

14.1

12.2

11.5

13.6

14.3 0.00018

20000

9.3

10.3

10.9

10.8

10.2 0.00018

Table 3 Measurement result of jitter on PWM drive system
Rotational
Displacement Δt
Jitter
speed
nsec
±%
min-1
500 101429.1 97723.3 77618.2 93284.7 89962.5 0.04226

Jitter (%) = ±{（Δt / T）×100 } / 2
-133-

1800

5664.65950.36932.75636.85379.7 0.01040

5400

685.8

752.9

531.4

582.7

781.4 0.00352

10000

237.5

191.5

220.1

357.4

268.6 0.00300

15000

128.3

117.8

163.6

135.4

149.5 0.00205

20000

162.3

72.8

121.3

184.2

97.2 0.00307

NTN TECHNICAL REVIEW No.79（2011）(2) Sealing performance

Certain foot components that support the air spindle
include a movable portion in order to accommodate
axial thermal displacement.
The air spindle was continuously run at a speed of
20,000 min-1. Fig. 9 plots the resultant variation on thesurface of the workpiece being machined. As
illustrated by this diagram, NTN’s newly developed air
spindle achieves a stable state within several minutes
after the start of operation, and regulates the variationon the surface of the workpiece to 1 µm or smaller.

Foreign matter trapped in the bearing of an air
spindle will jeopardize rotational performance and maylead to catastrophic bearing damage such as seizure.
Air spindles for machine tools may be used in severe
operating conditions where they are subjected tosplashes of cutting chips and coolant. Air spindles areequipped with a non-contact seal mechanism to protect
the bearing from these attacks. This seal mechanism
includes a narrow labyrinth gap between the rotary
shaft and fixed portion (housing); compressed air for
air-sealing is allowed to flow through this gap toprevent cutting chips and coolant from entering thebearing. To verify sealing performance of this sealing
system, the air spindle was continuously run for 24hours under coolant splash.

Variation on surface of workpiece being machinedμm[3] Improved shape

Water cooling: flow rate 2.5 L/min (setting at 23˚C)
Rotational speed: 20,000 min-1 continuous

1 μm max.

3.0
2.5(3) Chuck mechanism
For easy mounting/dismounting of the workpiece,
the air spindle is equipped with a vacuum chuckmechanism. A workpiece is firmly attracted to thechuck by the pressure difference relative toatmospheric pressure, because air between theworkpiece and table is evacuated from the oppositeto-table side through vent holes situated along theshaft core of the rotational shaft. The opposite-to-tableside is provided with a non-contact seal, drain trap
and drain port; this layout prevents the coolant fromentering the bearing (otherwise the coolant can be
drawn into the bearing through the air vent holes when
the work piece is chucked).

2.0
1.5
1.0
0.5
0
0

5

10

15

20

25

30

35

40

45

50

55

60

Time elapsed min

Fig. 9 Variation on surface of workpiece being machined

Drain port

Feed water port

Noncontact seal
Air vent holes
Table

Water jacket

Rotary spindle

Air seal
Air intake port
Workpiece

Drain port
Chuck
Air seal
Drain trap
Journal bearing

Drive motor

Thrust bearing
Compressed air supply for air seal

Fig. 10 Internal structure of NTN's air spindle for machine tool

-134-

EncoderAir Spindle for Ultra High-precision Machine Tools

6. Conclusion
NTN has performed performance evaluation tests
for its newly developed air spindle for ultra highprecision machine tools that meet contemporaryusers’ needs, and verified the commercial usefulnessof this product.
So far, NTN has designed and marketed optimalmechanical structure and control systems that meetthe cost and performance that users demand.
However, our engineering challenges moving forward
will be further complicated and more demanding. To
be able to promote users’ acceptance andpropagation of NTN products throughout its users,
continued engineering improvements for enhanced
performance, as well as timely performanceverification, will be increasingly important.
On the basis of its accumulated technologies andknowledge, NTN will further develop and improve its
air spindle product, and continue to seek new
applications for the products.

Reference
1) Horiuchi, T., Aono, K., Hiyoshi, H., Development of
Aerostatic Bearing Spindle for Precision Machine Tool,NTN Technical Review No. 74, P. 32-35, 2006.

Photo of authors

Teruyoshi HORIUCHI

Kazuyuki AONO

Precison Equipment Division
Product Engineering Dept

Precison Equipment Division
Product Engineering Dept

-135-

NTN TECHNICAL REVIEW No. 79（2011）

Presentation of award-winning articles
2010 ‘Cho’ Monodzukuri Innovative Parts andComponents Awards—Automotive Component Award

Torque Diode for Seat Lifter
Masahiro KAWAI

process(see Fig. 2). In 2002, for the first time webegan sales in Japan as a clutch for high-performancelever-operated manual seat lifters.

1. Introduction
This torque diode for seat lifters received theAutomotive Component Award at the 2010 'Cho'
Monodzukuri Innovative Parts and Components
Awards, which are sponsored by the Conference for
the Promotion of Monodzukuri and the Nikkan Kogyo
Shimbun (Business & Technology Daily News) withsupport from the Ministry of Economy, Trade andIndustry and The Japan Chamber of Commerce andIndustry.
The torque diode for seat lifters is used in themechanical device for manually adjusting the height ofa vehicle seat. The torque diode has the merits ofquiet operation, while allowing for invariableadjustment with only a small amount of force needed
to operate. It also contributes to making vehicle seats
more compact and more comfortable.

Fig. 2 Component parts

2. Structure of Torque Diode for Seat Lifters

3. Conclusion

The torque diode for seat lifters is a product thatwas developed jointly with Shiroki Corporation as a
new mechanical device that controls transmission of
torque between the input and output shafts.
Beginning with a wedge-shaped clutch structure(see Fig. 1) and by combining the high-precisionpress operations and lubrication techniques that NTN
has developed in its bearing technologies, NTN has
succeeded in creating a product that meets the needsof the market.
In order to meet the price requirements of themarket, the amount of cutting required was greatly
reduced by adopting a new high-precision press

At this time, vehicle seats with manual lifters are
growing at a rate of about 80% according to ourestimates. So we are contributing a great deal to the
popularity of the level-operated seat lifter models.
For now, we will advance the development of
products to achieve better performance with the goalof expanding in markets overseas.

Photo of author

Wedge

Masahiro KAWAI

Fig. 1 Wedge-shaped clutch (one-way clutch)(left: clutch engaged, right: clutch idle)

Automotive Business HQ
Axle Unit Engineering Dept.

-136-

NTN TECHNICAL REVIEW No. 79（2011）

Presentation of award-winning articles
First Certificate of Tribology Heritage

Journal Rolling Bearing for
the 0 “Zero” Series Shinkansen Trains
Kouya OHIRA*

Masanori UENO**

standards from that time have been contributing greatly
to the advancement of bearings for railway.
Shinkansen journal bearings are one key industrial
product that has continued to support the high speed,
safety and comfort of Shinkansen trains to this day.
These rolling bearings were recognized as animportant tribology heritage that made a great
contribution to the advancement of tribology history asa milestone for journal bearings in high speed trains,
leading to this designation.
Since the importance of the railway industry
continues to grow, there are demands for bearings thatare able to provide even higher speed, greater
reliability and longer life because trains offer highspeeds, high-capacity means of transportation with low
environmental impacts. We will continue to researchand develop journal bearings for railway that takesafety and economy into consideration while being
good for the environment.
This journal bearing model is being displayed in thefirst-floor lobby of our Industrial Technical Center in
Kuwana-shi, Mie.

Overview
At the 55th General Meeting of the JapaneseSociety of Tribologists, NTN’s journal rolling bearing for
the 0 Series Shinkansen trains was recognized as atribology heritage. 1) A system of honoring tribology
heritages was established this year with the goal ofidentifying actual products related to tribology that
contributed significantly to the advancement of scienceand technology and to preserve technologies.
These journal bearings were developed for the 0Series Shinkansen as Japan’s first high-speed train put
into operation when the Tokaido Shinkansen Line
commissioned in 1964. These journal bearings arecomprised of double-row cylindrical roller bearing andsingle-row deep-groove ball bearing. Journal bearings
bear the radial load, which is the weight of the train,
and the axial load, which occurs when traveling aroundcurves. The cylindrical roller bearings carry radial load,
while the deep-groove ball bearing bears the axial
load.
Moreover, in order to prevent the ball elements fromslipping on the raceways surface when no axial load ispresent, the deep groove ball bearing is preloaded
using a disc spring. Before Shinkansen, journalbearings on conventional trains were greaselubricated. In contrast, Shinkansen trains were the firstto adopt journal bearings lubricated by oil bathlubrication system because these bearings have tooperate at high speeds for a prolonged time.
As a result of journal bearing design development
and testing, a variety of specification and inspection
standards were established with the guidance of theJapanese National Railways (now JR). A variety of theReferences
1) Tribology Heritage #3: Journal Bearings for 0 Series
Shinkansen Trains, Journal of Japanese Society of
Tribologist, Vol. 56, No. 7, p. 402, 2011.

Fig. 2 Commendation from Japanese Society of Tribologist(From the left: Masayuki Mori, President, Japanese Society of
Tribologist, and Yosikazu Fukumura, Executive Director, NTN CORP.)

Fig. 1 Certificate of Tribology Heritage* Industrial Technical Center**Industrial Business Headquarters, Construction Machinery and Railway Engineering Department-137-

NTN TECHNICAL REVIEW No.79（2011）[ New Products Information ]Sealed Self-Aligning Roller Bearing(ULTAGE WA Type)With an additional seal that prevents the ingress of foreign matter,load capacity of the world’s highest level
Longer bearing life, higher reliability and easier handling!

Features

Structure[1] Worlds highest load capacity

High load capacity design• Adoption of the ULTAGE EA type internal design

Oil hole[2] Compact design that minimizes spacetaken up by seals in an assembly• Use of a special contact-type rubber seal design
prevents the ingress of contaminants.
• Contact pressure of the seal on the bearing
remains unchanged in spite of alignment operationon the bearing; the seal maintains its dustprevention performance during misalignment
conditions.
New contact-type seal

[3] Long-life grease used as standard

Applications¡Raw material transportation conveyors(in steel mills, thermal power generation stations, mines, etc.)¡Parts possibly attacked by salty air(transportation equipment in harbors; and construction
machinery, etc.)-138-

Prefilled grease:long life and high load
capacity

New Products Information

ULTAGE Sealed Four-Row Tapered Roller
Bearing for Rolling Mill Roll-Necks
Through optimized bearing design, we have greatly improved load resistance,load capacity and sealing performance, helping extend bearing replacement cycles
and improving reliability of the bearing during rolling mill operation!

Features (compared to previous models)

Structure[1] Worlds highest-load capacity design

Size reduction
in seal case

• Bearing rating life improved by as much as 1.8
times.

Seal of improved water
blocking performance[2] New crowning form improves load
distribution• Heavy load resistance more than doubled(world’s best)[3] Improved Seal Performance• Water resistance more than doubled

Applications

High-load capacity design Heavy load resistance design
Rollers and cage designed
for maximum load capacity¡ Steel mill roll-neck bearings-139-

Adoption of
new crowning form

NTN TECHNICAL REVIEW No.79（2011）

IC Tag Integrated Bearing
The industry’s first! Electronic data can be recorded inside the bearing itself.
Enhanced reliability about regular inspection information, and simplificationof otherwise time-consuming information management procedure!

.5

φ4

of applications and typical
Structure (examplescontentsof data being displayed))

Features[1] Confirmation of quality data reading canbe conducted from the bearing itself.
[2] Users can write and read usage history
to the bearing itself.
[3] Management items can be set freely.

[Image of use]

Applications¡Construction machinery, mining equipment,
aircraft, wind power generation facilities,
rolling stock, etc.

[Examples of data sets used for bearing life management](Bearing quality information)(suitable for applications that require regular maintenance)

1

Product name2

Bore dia., outside dia., precision of bearing width, gap information, etc.

3

Serial No.

4

Manufacture date, inspection date, shipping date

(User management information)

-140-

1

ID No. of machine that is equipped with the bearing in question

2

Date where the bearing is installed to the machine

3

Date where inspection takes place4

Accumulated operating hours

New Products Information

Large Bearing with Integrated Rotary Sensor
First in the industry!
NTN has created large bearings with incorporated rotary sensors!
Using this, a fewer number of parts and components are needed for installing a rotary
sensor and the time needed for installing and adjusting the sensor is virtually eliminatedFeatures

Structure (example: tapered roller bearing)[1] Rotary sensor is built into the bearing.
[2] A unique design has been adopted toprevent possible disconnection ofelectric wire between the shaft and innerring even if creeping (slipping) occurs
between the shaft and inner ring.
[3] Can be used immersed in oil.

Outer ring: as rotary ring

Pulser ringSensorWiring• Even in oil at 120˚C, the sensor reliably detects thebearing speed (can depend on oil type andadditives).
Inner ring: as fixed ring[4] Can reliably operate under vibrating
conditions.
• Even in severe vibrating conditions with vibration
acceleration of 10 G, the sensor reliably detects
the bearing speed.

Outer ring side
Pulser ring

Applications

Sensor¡Construction machinery, steelworks
facilities, wind power generation
facilities, ventilators, etc.

Inner ring side
Sensor-141-

Wiring

NTN TECHNICAL REVIEW No.79（2011）

New Standard Cam Follower
Compared to conventional cam followers, this product boasts
longer life and lower torque even in severe applications

Features (compared to previous models)

Structure[1] Longer rolling fatigue life achieved.

Non-relubrication
type• More than double the fatigue life whencompared to previous designs (with prefill ofhigh-performance grease)

Prefilled high-performancegrease[2] Non-relubrication type is prefilled withhigh-performance grease, boasting
greater wear resistance and much lower
starting torque at low temperatures.
• 35% reduction in starting torque (at -20˚C)[3] Installation and lubrication methods
remain the same as those of current
standard products(fully compatible)

Relubrication
type

Applications and series model numbers¡Transportation equipment, machine tools,
printing presses, press machines, etc.
¡NTN KR (V) 16–26 series models
With cages of outer ring outside diameters φ16–26
mm (stud diameter φ6–10 mm), or full complement
type roller cam followers-142-

Crowned rollers haveundergone special heat
treatment, increasing rolling
contact fatigue life

New Products Information

Rotary Sensor-Integrated Sliding Bearing
Combining resin sliding bearings and sensors helps simplify the installation of the
bearings into the machine. The design promotes reliability in monitoring of running state of
the machine, helping reduce time needed for maintenance work on the machine

Features

Structure[1] Unitized sliding bearing and rotary(magnetic) sensor(2) Sensor housing(1) Magnetic encoder• Easy handling and compact size

[2] Greater freedom in designing wiring
route

Radial bearing(sliding bearing)• Selectable orientation of the connector duringmounting.

[3] Wide input voltage range (3–26 V)• Compatible with diverse machinery andspecifications

Rotation
Bearing
detection part

Cutaway of bearing[4] Greater freedom for forms of bearingsand product outer diameters

D cut area
Retainer ring

• Customized design possible

Machine-side frame

Applications¡Rotary drives in office automation equipment,
measuring equipment, etc.
Rotary shaft
Sliding bearing withintegrated rotary sensor
Typical example of mounted bearing-143-

NTN TECHNICAL REVIEW No.79（2011）

Monodrive Two-way Feeder
with Spring Separation Mechanism
Enhances functionality of automatic spring parts feeding system.
Reduces necessary floor space while realizing reliable parts feeding!

Features

Structure

As automatic feeding system for tangled springs[1] Compact• Required floor space is about half (compared tobowl type models).
• Newly-developed spring separation mechanism
has been integrated into a compact monodrive
two-way feeder.

Tangledsprings
Separatedsprings[2] Stable supply• Tangled springs are rapidly and reliably separated
and transported by the separation mechanism.

Applications¡Automatic feeding equipment for small diameter
coil springs (diameters of up to 2–5 mm)
Spring separator-144-

New Products Information

Compact, Variable Frequency Controller
for Parts Feeders
Numerous functions are combined to contribute to a reduction in necessary
floor space and improvement in production efficiency

Features

Comparison of parts feeder product types[1] Reduced floor space(compared to previous models)

Newly-developed product:
K-ECF25

Conventional model:
K-ECA46• Volume: 1/3 (K-ECG25)–2/3 (K-ECH45)[2] Numerous functions• Resonance point tracking function (included inhigh-performance types)→ Up to 30% reduction in electricity consumption• Constant amplitude function (included in highperformance types)→ Transportation speed remains unchanged
regardless of variation in the volume of loaded
work pieces.
• Works depleted detection, and inching operation
functions (included in all types)
Compact bowl feeder: model K-K14

LineupNote: For details about applicable parts feeder unit being controlled, please refer to NTN’s relevant parts feeder catalog.

Model

Size mm

K-ECF25
W60×H140×D142
K-ECG25
K-ECH45
W82×H140×D142
K-ECJ45

Weight kg
About 1.1
About 1.2
About 1.7
About 1.7

Applicable units
K10–K20, N25, L20S, 05–S30, etc.
K20･1, N25–N40, G50-145-

Constant amplitude & resonance point tracking functions
No (standard type)
Yes (high-performance type)No (standard type)
Yes (high-performance type)