**表紙E_レイアウト 1 14/07/30 14:55 ページ 1

For New Technology NetworkR

TECHNICAL REVIEW No.81
Special Issue

corporation
http://www.ntn.co.jp

Special Issue;
Automotive Products andTechnologies
November 2013

Automotive Products and Technologies
corporation

This catalog is printed on recycled paper for environmental protection purposes.

ISSN 0915-0528 OSAKA, JAPAN

**表紙E_レイアウト 1 14/08/11 16:25 ページ 2

NTN to verify practicality of in-wheel motor driven vehicles throughthe demonstration project of microcompact mobility commuter EVs

NTN has started a demonstration project aiming for early expansion of electric vehicles that adopt an in-wheel motor
system, by creating prototype in-wheel motor driven microcompact mobility commuters.*
In Ise-City, Mie Prefecture, NTN is participating in the "Ise Municipal Council for the Creation of a Low-Carbon Society
Using Electric Vehicles, Etc.” and our EVs will be used as official cars of Ise-City, as well as business vehicles for theCouncil members such as Ise Chamber of Commerce and Industry, Ise City Tourist Association, etc. They are also used in the
tourism business partnering with the council members to verify their usefulness and impact on the environment.
In addition, NTN participated in the "EV Festa” which took place at Mount Fuji, which was recently listed as a World
Heritage site, as well as proceeded with the demonstration project on the public roads partnering with Iwata-City, Shizuoka
Prefecture. Overseas, NTN is partnering with Annecy, France to conduct demonstration projects by creating vehicles thatcomply with the European microcompact mobility vehicle standards together with the small to medium companies in the region.
Through these projects, NTN aims to expand the use of environmentally-friendly electric vehicles and verifying
performance of in-wheel motor driven vehicles characterized by their high degree of freedom for the space design and theiruse in real life.
*One- or two-passenger vehicles, with smaller scale and higher maneuverability than the“light cars,” for neighborhood commuting with high environmental performance▲Ceremony of lending vehicles at Ise-City(September 30, 2013)(Center: Ise-City Mayor Suzuki
Right: Council Chairman Paku
Left: NTN Chairman Suzuki)

NTN TECHNICAL REVIEW No.81
Printed and published in August 29, 2014
▲Mt. Fuji EV Festa (August 1, 2013)(Right: Shizuoka Prefecture Governor Kawakatsu
Left: Association for the Promotion of Electric Vehicles,
Chief Commissioner Tajima)

EditorYukiyoshi HAGIMOTO

Publisher Koji TAKAHASHI
Published by NTN Corporation
3-17, 1-chome, kyomachibori, Nishi-ku, Osaka,J apan

At the opening event of
Annecy International Animated Film Festival(June 10, 2013)

Printed by
NISSHA Co., LtdNo part of this publication may be reprinted or reproduced
without the prior permission of the copyright owner.

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R

TECHNICAL REVIEW
No.81

Special Issue:
Automotive Products and Technologies

corporation

*01-目次-巻頭言E_*01/目次/巻頭言 14/07/31 10:53 ページ 2

NTN TECHNICAL REVIEW No.81
CONTENTS
Preface
Contribution

Yukiyoshi HAGIMOTO

Trends of Intelligent Vehicle Dynamics Controls and Their Future
Masato ABE Professor, Department of Vehicle System Engineering, Kanagawa Institute of Technology

Perspective

NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles
Takehiko UMEMOTO

1

2

12

● Special Issue for EV/HEV Products
Recent Technology Trends of In-Wheel Motor System for Automotive
Tomoaki MAKINO, Aiko ISHIKAWA, Chinami ITOU and Kayo SAKAIElectro-mechanical Brake Unit with Parking Brake
Tatsuya YAMASAKI, Makoto MURAMATSU and Yui MASUDA

Ball Screw Drive Module for Electric Hydraulic Brake

22
30

Hayato KAWAGUCHI and Keisuke KAZUNO

36

Low Torque Deep Groove Ball Bearings for EV and HEV

40

Tsuyoshi KODA, Takahiro WAKUDA and Tomohisa UOZUMIThe Reactor Core for HEV Boost Converter
Takayuki ODA, Shinji MIYAZAKI and Eiichirou SHIMAZU

46● Special Issue for Advanced Function and Fuel-efficient Products
Hub Bearing with an Integrated High-Resolution Rotation Sensor
Kentaro NISHIKAWA, Toru TAKAHASHI and Christophe DURET

52

Press Connect Spline Hub Joint

58

Takayuki NORIMATSU and Tsutomu NAGATA

Next-generation High Efficiency Fixed Type Constant Velocity Joint "CFJ"Teruaki FUJIO

Development of Low Torque Seal Ring for Automotive Transmission
Kouzou KAKEHI, Takumi KONDOU, Takuya ISHII and Masato YOSHINOThe Introduction of Sintered New Products for Automobile
Tomonori YAMASHITA, Tomokazu SONOZAKI

64
68
74

● Technical Papers
Rolling Contact Fatigue Life of Highly-carbonitrided JIS-SUJ2
Daisuke SATO and Chikara OHKI

Effect of Naturally Derived Antioxidants as Surface Treatment
Yosuke TAGUCHI, Eriko ITAHASHI and Hidenobu MIKAMI

Grease Leakage Prevention for Double-Row Angular Ball Bearings

78
85

Mitsuo KAWAMURA, Norihide SATO, Tomoya SAKAGUCHI and Mariko IZUMI92

A Creep Mechanism of Rolling Bearings

100

Tsuyoshi NIWA

● Our Line of Award Winning Products
The Japanese Society of Tribologists "2012 Technology Award"

Prevention of Grease Leakage from Ball Bearings Due to Cage Shape Improvement

104

Norihide SATO, Mitsuo KAWAMURA and Tomoya SAKAGUCHII
European Mechatronics Meeting (EMM) 2012 "Grand Prix of Mechatronics Award"Intelligent In-wheel Motor System

105

Herve LENON, Eiichi ASANO, Tomomi GOTOU, and Kaori TOMODA
Automechanika 2012 "Innovation Award"

In-wheel Motor System

106

Herve LENON and Eiichi ASANO,The Japan Society of Precision Engineering of the 2013 Spring Semestrial Meeting "Best Presentation Award"

Parallel Link High Speed Angle Control Equipment

107

Hiroshi ISOBE, Yukihiro NISHIO, Keisuke SONE, Hiroyuki YAMADA and Yoshio FUJIKAWA

Our Line of New Products108

*01-目次-巻頭言E_*01/目次/巻頭言 14/07/31 10:53 ページ 3

NTN TECHNICAL REVIEW No.81（2013）[ Preface ]A message for the special issue on
automotive products and technologies

Yukiyoshi Hagimoto
Managing Executive Officer

As the motorization in developing countries is fully in progress, the automobile industry isfacing the critical challenge of low-fuel consumption requirements for reduction ofgreenhouse gas emission and the use of fossil fuels such as oil. Therefore, an aggressive
development of engines for lower fuel consumption, hybrid vehicles and electric vehicles isunderway. Also, development of fuel cell vehicles is becoming active with the target ofcommercial introduction in 2015.
In view of these circumstances, NTN reorganized our engineering functions within the
Automotive Business HQ to accelerate the development of unit and module products to
support energy conservation of automobiles, and established a new EV Module Business HQ
in April, 2013 to promote development of parts, modules and systems related to the driving,steering and controlling of electric vehicles.
We have worked to publish this issue of the NTN Technical Review No. 81 featuring thestate-of-art automotive products and technologies in time with many exhibits of new products
from NTN at the 43rd Tokyo Motor Show 2013, which was held during the period from
November 22 to December 1, 2013 with the theme of "Compete! And shape a new future." In
this Technical Review, we were able to publish a contribution “Trends of Intelligent Vehicle
Dynamics Controls and Their Future” from Professor Masato Abe of Department of Vehicle
System Engineering, Kanagawa Institute of Technology, who is an expert in vehicle dynamics
control. In the “Special Issue for EV/HEV Products," we introduce NTN’s driving module,
control system, etc. which lead to automatic driving including an in-wheel motor system forelectric vehicles. In the "Special Issue for Advanced Function and Fuel-efficient Products," weintroduce hub bearings and constant velocity joints which NTN has been globally distributing
and new products contributing to higher functionality and lower-fuel economy, as well ascomposite material products consisting of mainly resin and sintered material.
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 contribute to sustainable
development of the society by further developing environmentally-friendly products.

-1-

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NTN TECHNICAL REVIEW No.81（2013）[ Contribution ]Trends of Intelligent Vehicle Dynamics Controls
and Their Future

Masato ABE
Doctor of Engineering, Professor, Department of Vehicle System Engineering,
Kanagawa Institute of Technology

Following to a survey of recent trends of active motion control of the vehicles, a possibility of thevehicle active motion control of electric powered vehicles is discussed and the experimental
substantiation of the possibility by using the experimental vehicle is introduced. It is found that thevehicle motion controls which are not available for the ordinary vehicles so far will become available byan ultimate intelligent active motion control of electric powered vehicles. A vehicle limit performance isdramatically extended and the tire wear of the vehicle during controlled vehicle motion is significantly
reduced by the controls.

equipped in commercial vehicles that are greatly
improving their safety and mobility performance. It is
expected that active motion control based on
intelligent technology will increase among others,
since the motion control which has been impracticalbefore now becomes more realistic as electric
powered vehicles expand.

1. Introduction
Along with the adoption of the electric motor as oneof the pillars of the next-generation vehicles to copewith environmental concerns and energy issues,
another significant technology area supporting next
generation vehicles is the so-called intelligent/intellectual
vehicles. Fig. 1 shows a representation of technology
areas supporting next-generation vehicles in one ofthe views. In the broad "Environment andTransportation System," the next-generation vehicle
can be defined as electric "power" mounted on the"structure" for "motion" and it seems like one of the
key technologies in this picture is the "intelligence" forvehicles. The specific technology elements in this area
include car electronics, information, communication,
control, etc.
Intelligence for vehicles is the technology thatshould be related to almost all automotive
technologies, however, especially in the active control
area aiming at prevention and safety, it is not only an
indispensable technology for motion control ofindividual vehicles but also for mutual motion control
of several vehicles in the transportation space.
Several active motion control technologies are alreadyIntelligence

Car electronics
Control/information/communication area

PowerElectric vehicle system
area
Electric vehicle
Hybrid vehicle
Fuel cell vehicle

ITS

MotionAdvanced vehicle motion
performance area

Structure

Active control
Drive assistancePreventive safety

Smart structure area
Light-weight/highly
efficient structure
Collision safe

Environment / ransportation
system

Fig.1 Technologies supporting a vehicle for new era-2-

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Trends of Intelligent Vehicle Dynamics Controls and Their Futureobtained from the distribution of driving force on theright and left sides, distribution of braking force orbrake on only one side, etc. This type of motion
control by yaw-moment may be seen as existing for along time like the passive control by differential gears,
however, it was only since the 1990s that many
research and development activities for active motion
control by actively producing yaw-moment wereobserved.

In this article, I will give an overview of the active
motion control that has been proposed andincorporated so far within the horizontal plane which adriver is directly involved in the control and is largely
related to the preventive safety from the view point of
the motion dynamics property of controlled vehicles,
as well as some specifics on the trends andpossibilities of active motion controls in a new era ofincreased use of electric power in vehicles. I will also
describe some results from our ongoing joint researchwith NTN.

The lateral force generated by the side slip isgreatly affected by the lateral motion, load, andlongitudinal force of the vehicle to be controlled. In
contrast, the longitudinal force is not fundamentallyaffected by the lateral motion of a vehicle. Some may
argue that the longitudinal force should also beaffected by the load and lateral force or lateral motion.
However, the command of longitudinal force
necessary for control is the braking torque applied bybraking pressure and driving torque given to the tire
rotating motion. The tire rotation motion itself is
independent from the lateral motion of a vehicle andthe tires follow the command torque providing rotation
motion. The longitudinal slip produced by this motion
generates the longitudinal force that balances with thecommand torque. Therefore, the longitudinal force is
not directly affected by the lateral motion and cancontrol the lateral motion with the required force. That
is of course limited to the cases when there issufficient tire load to produce a longitudinal force large
enough for the corresponding command. This appears
to be the technical characteristic at the base of thefact that the electric stability control (ESC), one type of
DYC, is about to be standardized and required, in
addition to the fact that deceleration is effective inorder to stabilize the motion.
The vehicle motion on the horizontal plane against
steering and external yaw moments can be expressed
essentially by the two-variable motion equation oflateral slip and yaw. By looking at this equation, it iseasily understood that when a driver steers the front
wheels the active motion control that exhibits thesame motion as that following the steering can befundamentally achieved whether it is by the rear-wheel
active steering or by DYC. Fig. 2 shows thecomparison of the results from the rear-wheel active
control and DYC when the yaw rate is controlled to be
the response to the primary delay of steering1). Thisshows the motion in terms of side slip and yaw rate
locus from one cycle of the sine wave steering input.
From the graph we can compare the magnitude of theside slip produced during the sine wave yaw motion.
The graph shows the impact of disturbance of
longitudinal acceleration and deceleration during the

2. Past active motion control technologies
2. 1 Rear-wheel active steering
In the 1980s, a breakthrough technology of rearwheel active steering control made its debut in theimpasse of vehicle motion performance. It had someprecedence before; however, this is a period whenmotion control by rear-wheel active steering wasviewed as an improved motion control in a relatively
high-speed range. Initially, rear-wheel steering relative
to steering angle and force was proposed to improve
motion property, for example, eliminating or reducing
the side slip to the maximum extent. Then, more
intelligent model-following control systems were
proposed such as the feed-forward control in which
the yaw motion of a vehicle follows the model
response and addition of feedback control following
the detection of the yaw rate. With these technologicalresults in the background, commercial vehicles withrear-wheel active steering control were introduced in
the market.
The rear-wheel steering control depends on thecornering force of the tires generated by the side-kickangle from steering. For example, the control law
driven from the linear model with the two-degrees-offreedom model is effective in the range where theside-slip angle is small exhibiting a linear tire property.
However, in the non-linear range of lateral force of thetire, the aimed control ability is drastically reduced. In
addition, since the lateral force of the tire is greatly
affected by the load and longitudinal force theaccuracy of motion control can be lost when the lateral
acceleration is large and furthermore, the motion
involving acceleration and deceleration can become adisturbance.

2. 2 DYC
The technology that emerged next was the direct
yaw-moment control (DYC). DYC is a global name forthe motion control taking advantage of the moment
produced by the longitudinal force on the individual
tires. The longitudinal force for this technology can be-3-

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NTN TECHNICAL REVIEW No.81（2013）

Initial Vehicle Speed 100km/h
Single Sine Wave Steering Input
Without Control

0

30
0

-30

Feed-foward DYC

Yaw Rate
Feed-back DYC

-6

-6

0

6

-30

-6

0

6

Sideslip Angle [deg]Yaw rate [deg/s]

Yaw rate [deg/s]

30

Yaw rate [deg/s]Longitudinal Acceleration-0.45〜0.30 [G]

30

0

-30

Feed-foward 4WS

-6

0

6

0

steering is becoming electric power steering (EPS),
making it easy to control the steering reaction force. It
has been determined that 50% of the sensory
evaluation of a driver regarding the steering response
can be controlled by adjusting the steering reaction
force without changing any vehicle motion
performance. Thus, setting of the steering reaction
force is important, even if it does not directly
contribute to the active motion control; not to beignored in the overall scheme of motion control ofintelligent vehicles, and therefore, attracting muchattention.

6

Yaw Rate
Feed-back 4WS

-6

0

6

Sideslip Angle [deg]2. 4 Other control

Fig.2 Effects of rear wheel active steer and DYC

Active suspension, which is motion control of avehicle with respect to comfort, is outside the scope of
this article; however, the active stabilizer in thattechnology is used for controlling the lateral motion of
a vehicle. This technology does not directly use thelateral force; however, it controls the load on the rightand left tires which affect the lateral force and as aresult controls the lateral motion.
Other important controls worth mentioning, although
they do not directly control the motion of a vehicle
itself, are ABS and TCS. They control components
which control the functions of tires that support vehicle
motion and as a result greatly contribute to theimprovement of the vehicle motion performance.

motion indicating that the control by DYC is morerobust against the disturbance with more stablemotion even in the non-linear range of large side slipof tires. However, in case of rear-wheel active
steering, it is revealed that an adverse effect was
produced due to the limitation of the tire against lateralforce, even with the yaw rate feedback trying to force
it to follow the model response.
Therefore, DYC is more superior to the rear-wheel
active steering in many respects. Although DYC is
suitable for directly controlling yaw motion, it is notnecessarily suitable for controlling the side slip angle
and lateral acceleration since it does not directly
produce lateral force 2).

2. 5 Integrated control
Even with acceleration and deceleration, if thelongitudinal speed is constant or given as a parameterthat changes during its period, the vehicle motion in
the horizontal plane against steering can be basically
expressed with two-degrees of freedom of side slip
and yaw. By using any two of the aforementioned
controls, namely, front-wheel active steering, rearwheel active steering or DYC, a vehicle that provides
any side slip and yaw motion against steering can berealized theoretically. One idea of the integrated
control is to obtain desirable motion properties byusing two (or more) active controls simultaneously
based on this perspective.
Each active control seems to have its own limitation
if used individually, as mentioned before. The
limitation perhaps comes from the non-linearity of tire
properties, as some extreme load may be placed ontires during the control. When two or more controls areused, this can be offset by a synergy effect thatcompensates the limitations with each other. The
integrated control can now be described as using two
or more controls simultaneously and improving thevehicle motion properties and their limitations by using
4-tire forces reasonably 3), 4), 5).

2. 3 Front-wheel active control and steering
control
A motion control with further active steering control
over a driver’s steering of the front wheels is also
being proposed and realized. It appears to have thesame problem as the rear-wheel active steering as it
depends on the lateral force from the steering angle,
however, the mechanism for realizing this method canbe simpler than reaching out to the rear wheels forsteering. In addition, it is possible to obtain theoreticaladvantages and disadvantages of the respective
active control methods from the motion equation ortransfer function of the side slip in the horizontal plane
and the response of the yaw motion to the front-wheel
or rear-wheel steering. The front-wheel active control
eliminating the mechanical link between front wheels
and steering wheel is a complete steer-by-wire (SBW)system. It appears that this system will be achieved in
the near future. This system increases the degree of
freedom for control and steering reaction force, among
others, and can be set at any value.
On the other hand, most of the current power-4-

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Trends of Intelligent Vehicle Dynamics Controls and Their Future

Fig. 3 is a comparison of the effects of the rearwheel active control, roll-moment distribution control(RDC) using a longitudinal stabilizer and theirintegrated control with side slip, and yaw rate locus
against a sine wave steering input 6). Rear-wheel
active steering is a front–wheel steering angle
proportional control to eliminate the steady side slip
and RDC is a control for the stabilizer rigidity of thefront and rear wheels maintaining the total roll rigidity
so that side slip of front and rear wheels match thedriver’s intention. The motion is stable if a constant
yaw motion is produced from a small side slip. With
only the rear-wheel active steering, the motion is notstabilized when the burden on the rear wheels
becomes large at high speeds and the side slip
becomes larger than without control. In contrast, when
it is used with RDC, the excessive transition of load
between the left and right rear wheels is suppressed
resulting in improvement of lateral force limitations
realizing the motion aimed by the rear-wheel active
steering up to the high speed range. This effect islarger than what can be obtained with RDC only.
There are already some examples of commerciallyavailable vehicles with integrated control. If thecapabilities of four tires are more reasonably used, it
appears that the vehicle motion properties andlimitation performance can be further improved. From
this standpoint, the integrated control is likely toremain as one of the most important key words invehicle motion control.

-8.0

-4.0

20

120km/h
100km/h

10
4.0

8.0

-8.0

-4.0

-10

-10

-20

-20

-4.0

3. 2 Distribution control and norm

4WS

Yaw rate [deg/s]

140km/hYaw rate [deg/s]

20
10

-8.0

RDC Only

30

120km/h
100km/h

120km/h
100km/h

4.0

8.0

-30

-30

Without Control
4WS

4.0

Sideslip Angle [deg]Sideslip Angle [deg]140km/hRecently, there has been an emergence of electricpowered vehicles. The first motivation for this trend, ofcourse, is to improve environmental/energy concerns.
However, if those electric-powered vehicles areequipped with four in-wheel motors, the longitudinal
force of these four tires can be freely andindependently determined. Also, since there will be no
drive shaft, these four wheels can be easily steered
independently and the lateral force can be also freely
and independently determined. This means that eightforces that determine the motion of a vehicle can be
independently determined. Of course, these vehicles
are equipped with "full drive by wire." From thestandpoint of the dynamics of a vehicle, this presents
a tremendous possibility.
Although they can be freely and independently
determined, in order for a vehicle to respond in apredetermined way to the steering in a plane certain
lateral force and yaw moments need to be applied ateach instance to the vehicle. These forces can beobtained reversely if the target response is given. In
order to provide motion corresponding to the driver’s
acceleration or braking operation, longitudinal force isneeded based on those commands. The grand sum of
the lateral and longitudinal forces from those 8 forces
and the sum of the yaw moments produced by those 8
forces need to be always maintained at varying values,
determined by the target response of the vehicle ateach instance. This constitutes three constrained
conditions for 8 freely determinable variables.
On the other hand, since the forces can be
determined in this manner, any motion can bebasically achieved in response to the steering. When
only desirable response of the side slip and yaw rate
to the steering is determined, active motion control
that delivers desirable targeted vehicle motion can beachieved.

Yaw rate [deg/s]

10140km/hYaw rate [deg/s]

20120km/h
100km/h

3. 1 Eight-variable control

30

30
140km/h

3. Electric power and intelligent motion
control

8.0

Sideslip Angle [deg]

-8.0

-4.0

4.0

Since there are three constrained conditions to eightvariables, the redundancy of degrees-of-freedom isstill 5. Therefore, all eight variables cannot be
determined without some norm for determining them,which is the norm for how to use these eight tireforces. Any norm will work; however, one that can beeasily assumed is the norm (1) to use the tires so thatthe load on each tire is minimized as a whole. Thisway it is expected that the burden on each tire can beevenly distributed and the motion limit performancecan be improved.

8.0

Sideslip Angle [deg]

-10
-20
-30

Rear Wheel Steer Only

Rear Wheel Steer and RDC

Fig.3 Integrated control of rear wheel active steer
and RDC

-5-

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NTN TECHNICAL REVIEW No.81（2013）
Sharing X, Y & M with eight variables of four
wheels to minimize the cost function J

Another possible norm is (2) to control the motion
using the tire force so that the energy dissipation due
to the tire slip during motion is minimized on thewhole. As a result, it is expected to realize even
distribution of tire wear due to the tire slip and motion
control with minimal energy dissipation due to the tire
slip. This will minimize wear of tires.
If the evaluation function of the norm (1) is a squaresum of the resultant force from all the tires in thehorizontal plane divided by the tire load, it can be
expressed by the quadratic form of the distributed
longitudinal and lateral force of each tire. Similarly, if
the evaluation function of the norm (2) is a squaresum of the energy dissipation rate due to slippage ofeach tire, it can be expressed as the square sum of
the distributed longitudinal and lateral force of eachtire. However, each factor in these equations has tobe estimated in real time from the slip force and slip
velocity of each tire in motion using the tire model.
In either norm case (1) or (2), the evaluation
function can be given with the quadratic form of theforce applied to each tire with 8 variables; therefore,
the condition to minimize this under the threeconstrained conditions is given with a linear algebraic
function of 8-3=5 variables which can be solved quite
easily. In this way, these eight variables are finallydetermined and commands for driving or braking andsteering can be issued to each tire so that each tirecan produce these forces. This is the distribution
control of eight forces. Fig. 4 is a graphicalrepresentation of this distribution.
Given above, three constrained conditions wereminimally required and were applied against eightvariables that realize the targeted motion. Since it stillhas 5-degrees of freedom of redundancy, there isroom for additional constrained conditions to realize
some desirable motion. It is well known that thelongitudinal force working on tires can produce verticalforce to a vehicle depending on the suspension
setting. Therefore, a condition to produce rolling
moment to suppress vehicle rolling from this verticalforce can be set in the equation. If this is given as the
fourth constrained condition, the previous condition tominimize the evaluation function given in quadratic
form can be reduced to the linear equation of 4variables which can be solved even more easily. As
such, new constrained conditions can be added todetermine the distribution of eight forces to realize amore desirable motion. By taking advantage of theredundancy of the variables, motion can be controlled
by adding motion constrained conditions in therealistic range.

X, Y & M
Y1
Y3
X1

X3

8 variables !!
x
Y4

Y2
X4

X2

then, Ti and δi to steering inputδh are determined.
Y1
Y3
X1

X3

δ3

T3

δ1

T1
Y2

Y4
X4

δ4

δh

X2

δ2

T4
T2

Fig.4 Concept of 8 tire force distribution

3. 3 Estimation of tire energy dissipation
When an appropriate tire model is used, not only
the longitudinal and lateral forces on the tires can be
estimated real-time and online but also slip force andslip velocity. The energy dissipation rate due toslippage can be estimated from the slip velocity andslip force and its integration gives the tire energy
dissipation due to slippage.
For example, by using the brush tire model, which isoften used as a physical model, longitudinal force andlateral force in the adhesion region and slip region of
the contact area can be given by an analyticalequation when longitudinal slip rate, lateral slip angle,
traveling speed, and load are input. Similarly,
longitudinal and lateral slip velocity can also be given.
Therefore, a good estimate of tire forces can now be
obtained if the parameters of that physical model are
adjusted to the actual measurement.
This online estimation based on the model can beused for distribution control of the norm (2) and also
for the evaluation of energy dissipation due to tire slip
in motion.

-6-

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Trends of Intelligent Vehicle Dynamics Controls and Their Future

3. 4 Ultimate intelligent motion control and its
effect(1) Effect of distribution control minimizing tire
workload

Fig. 5 shows the block diagram of active motion
control using the distribution of eight forces to eachtire based on the aforementioned approach. Researchand development of intelligent motion control taking
advantage of the properties of electric vehicles bycontrolling these eight forces has been started indifferent parts of the world 7), 8), 9). It appears that theactive motion control of electric vehicles by "full driveby wire" will lead to the ultimate integrated active
motion controlled intelligent vehicles.
We also verified that this is realizable by using an
actual vehicle in the joint research conducted with
NTN, to demonstrate that this is not merely a desk
plan 10), 11), 12). Fig. 6 shows the experimental vehicle
for this demonstration with an in-wheel motor
equipped at each of four wheels and steering system
to independently control all four wheels. In addition, amicro-auto box in which the intelligent active motion
control system depicted in Fig. 5 is installed ismounted along with the vehicle motion measurement
system in the vehicle.

As mentioned earlier, any response to the steering isbasically possible in this vehicle motion control;
however, in order to compare the effect of the control
with the regular vehicle without control, the response to
be realized is set to regular vehicle response which canbe obtained from a two-degrees of freedom model oflateral slip and yaw on a plane. A braking lane change
experiment of -0.4G of longitudinal acceleration was
conducted on the course shown in Fig. 7 in order toconfirm the effect of distribution control of the norm (1)mentioned above under this setting.
The result is shown in Fig. 8. When no control isapplied, particularly in the latter half of the lanechange when counter steering is made, the motion
became unstable, possibly because the workload on acertain rear tire was extremely large exceeding thelimit. In contrast, when control is applied, the motion
was stable in all the cases of lane changes possibly
due to the load to the tires in motion was evenly
distributed without any specific tire bearing extreme
burden.

YawmomentDistribution 4-wheel drive
angle
control steeringcommand

VehicleVehiclemotion50m

2.5m

Slip force
Slip velocity

Tire
model

21, 17, 15, 12.5 m

50m

2.5m
Tire energy
dissipation

Fig.7 Lane change course

Fig.5 Tire force distribution control

Without distribution control

120

4

90
60

3

0

2

-60

1

30
-30
-90
-120

0

1

2

3

4

5

0

Lateral Displacement [m]Lateraforce

Steering Angle [deg]Distribution control minimizing tire workload

Fig.6 Experimental vehicle

120

4

60

3

0

2

-60

1

90
30

-30
-90
-120

0

1

2

Time [s]

3

4

5

0

Lateral Displacement [m]Accelerator
Brake

Targetvehicleresponse

Lane change length

4-wheel drive
brakecommandLongitudinal
force

Steering Angle [deg]Steering
angle

Fig.8 Effects of distribution control minimizing tire workload-7-

*02-寄稿文E_*02/寄稿文 14/07/31 10:16 ページ 7

NTN TECHNICAL REVIEW No.81（2013）(2) Effects of distribution control minimizing tireenergy dissipation due to slippage

The effect of control for even distribution of tire load
is significant. Attempting to control the vehicle rolling
by applying lateral and longitudinal forces in thereverse direction and of the same magnitude, themotion may become unstable due to imbalance of tireloads even if the rolling may be controlled. The normof distribution control for load equalization should
become effective, particularly in this case.
To confirm this assumption, a braking lane change
simulation of -0.4G from 120 km/h in a small
passenger car with a man-vehicle system was madeusing an appropriate driver model. The result is shownin Fig. 9. As anticipated, stable vehicle motion withsmall rolling was obtained only with "load equalization
distribution control + rolling control" where constrained
condition of rolling control was added. I am also willingto verify this with a real vehicle.

5

40

4

20

3

0

2

-20

1

-40
-60

0

2

0

2

4

6

4

6

Lateral Displacement [m]Steering Angle [deg]60

Using the same course, experiments of lanechanges were conducted including the distribution
control vehicle for the above norm (2) to examine theimpact of these distribution controls on the energy
dissipation due to tire slip during the motion.
Fig. 10 shows the vehicle motion when a lanechange with peak lateral acceleration of around 0.4G
is made at the constant speed of 60 km/h. From thegraph it is apparent that the motion is almost the samewith or without control. Fig. 11 shows how the energy
dissipation due to each tire slip and total energy
dissipation of all the tires change depending on thetype of control from the start of lane change to theend. In addition, Fig. 12 shows the motion of thevehicle when a steep lane change with peak lateral
acceleration of 0.6G is applied and Fig. 13 shows thetire energy dissipation during that time.
When the peak lateral acceleration with the lanechange increases from 0.4G to 0.6G, the energy
dissipation due to tire slip drastically increases;
however, the norm (2), which is the distribution control
minimizing the square sum of energy dissipation due
to tire slip in real time, is most effective for reducing
the energy dissipation due to slip in all cases. With
this control, among others, the energy dissipation ateach tire is equalized showing similar energy
dissipation at all the tires. It is expected that the tire
wear due to motion will be equalized with four wheelsby this control.
Although the tire load equalization control was
initially considered to be also effective for energy
dissipation, the result was adverse; showing larger
total tire energy dissipation due to slippage even whencompared to the case without control. Since theworkload ratio on each tire is equalized, thedistribution control produces a large force in thehorizontal plane at the tire with heavy load resulting inlarge energy dissipation due to tire slip by that large
force which contributes to large total energy
dissipation even compared with the case without
control. Even if this control contributes to stabilizing
the motion, it should promote imbalance of tire wear.
A control that distributes the tire force in real time andestimates the energy dissipation of each tire at each
instance due to slip based on a tire model is requiredin order to realize motion that minimizes tire energy
dissipation and tire wear.

0

Roll Angle [deg]

2
1
0
-1
-2

Without control
Roll control only

Time [s]Distribution control only
Roll and distribution control

Fig.9 Combined effects of both distribution
and roll controls-8-*02-寄稿文E_*02/寄稿文 14/07/31 10:16 ページ 8

Trends of Intelligent Vehicle Dynamics Controls and Their Future

-20

1

-40

3

20
0

2

-20

1

-40
-60

0.8

70

0.8

70

0.6

60

0.6

60

2

3

4

0.4

50

0.2

40

0

30

-0.2

20

-0.4
-0.6

10

-0.8

0

0

1

2

3

Time [s]4Lateral Acceleration [m/s2]

0

1

0

0

1

40
30

-0.2
-0.4

20

-0.6

10
0

1

Time [s]340Fig.12 Lane change with peak lateral acceleration 0.6G

Without control

Exact energy squaresum control

Tire workroad control
180

5

600

160

35
30

3

25
20

2

15
10

Dissipation energy of each tire [J]

4

1

500

140
120

400

100
300

80

200

60
40

Total tire energy dissipation [J]40Total tire energy dissipation [J]Dissipation energy of each tire [J]2Without control
Tire workroad control
Exact energy square sum control

45

100

20

5
0

0

50

0

Exact energy squaresum control

50

4

0.2

-0.8

Fig.10 Lane change with peak lateral acceleration 0.4G

Tire workroad control

3

0.4

Tire workroad control
Without control
Exact energy square sum control

Without control

2

Lateral Displacement [m]

2

0

40

Vehicle Speed [km/h]

20

Steering Angle [deg]

3

4

60

Lateral Displacement [m]

40

-60

Lateral Acceleration [m/s2]

4

Vehicle Speed [km/h]

Steering Angle [deg]

60

FL

FR

RL

RR

TOTAL

0

0

Fig.11 Effects of distribution controls on tire energy
dissipation (0.4G lane change)

FL

FR

RL

RR

TOTAL

0

Fig.13 Effects of distribution controls on tire energy
dissipation (0.6G lane change)-9-*02-寄稿文E_*02/寄稿文 14/07/31 10:16 ページ 9

NTN TECHNICAL REVIEW No.81（2013）

how to set the steering torque as well as how to apply
the method using this driver model and should also be
one of the important challenges for the ultimate
intelligent integrated active motion control vehicles.

4. Challenges
I have given an overview of the past active control
approaches for vehicle motion and described anultimate intelligent integrated active motion control
vehicle for electric vehicles and indicated that it is not
merely a desk theory.
Although this approach can basically realize any
response to the steering, the method that can begenerally discussed to determine which response is tobe realized is not yet proposed or established. This isan old issue about the relation between a driver’s
subjective evaluation of the vehicle steering properties
and objective vehicle motion performance. This
indicates that the method that depends on the driver’s
feeling is not reduced to the general method forvehicular motion performance design. The vehicle
should be easily maneuverable and highly perceived
by a driver and the vehicular motion performance
design should contribute to safety and also be "fun todrive."We have recently proposed a method to evaluate
vehicular motion performance using a driver model ofsimple steering angle control by identifying key
parameters of the model from the actual driver’s
behavior and vehicle motion 13), 14). The driverparameters identified in this method reflects thevehicle motion performance and this method should
provide a reasonable logic between the objective
motion performance and the driver’s evaluation. We
have seen many examples where the evaluation of
motion performance from this method matches withthe driver’s sensory evaluation very well. One of theongoing challenges should be to explore which
response to the steering that must be realized by theultimate vehicle active motion control based on, forexample, this method.
Also, when "steer by wire" or "drive by wire" isequipped, how to set the steering reaction will become
an important issue, as briefly mentioned earlier.
Currently, there is no generally established method.
The steering reaction is not an issue directly related to
the vehicle motion performance. Rather, it is an issueof the driver's evaluation. Although it is not an issue ofthe motion performance itself, in our recent research
we found that by changing the steering torqueproperty in the steering angle control driver model
mentioned earlier without changing any parameters in
the model for vehicle motion properties at all, themodel parameters would change in concert with thesensory evaluation as if the vehicle motion
performance had changed 15). It is quite interesting inReference
1) Y. Furukawa and M. Abe “Advanced chassis control
systems for vehicle handling and active safety”
Vehicle System Dynamics, Vol. 28, pp. 59-86, 1997.
2) M. Abe “Vehicle dynamics and control for improving
handling and active safety: from four-wheel steering
to direct yaw moment control” In proceeding of
IMechE, Vol. 213, part K, pp. 87-101, 1999.
3) G. Mastinu, E. Babbel, P. Lugner, D Margolis, D.
Mittermayr and B. Richiter “Integrated controls of
lateral vehicle dynamics” Vehicle System Dynamics,
Supplement 23, Vol.23, pp.358-377, 1994.
4) S. Sato, H. Inoue, M. Tabata and S. Inagaki“Integrated chassis control system for improved
vehicle dynamics” In proceedings of the AVEC92,
pp.413-418, 1992.
5) T. Gordon, M. Howell and F. Brandao “Integrated
control methodologies for road vehicles” Vehicle
System Dynamics, Vol.40, pp157-187, 2003.
6) M. Abe “A Study on Effects of Roll Moment
Distribution Control in Active Suspension on
Improvement of Limit Performance of Vehicle
Handling” JSAE Review Vol.12, No.3, July 1991
7) O. Mokhiamar and M. Abe “Simultaneous optimal
distribution of lateral and longitudinal tire forces forthe model following control” Transactions of the
ASME, Journal of Dynamic System, Measurement
and Control, Vol.126, pp753-763, December 2004
8) P. Reinold, V. Nachtigal and A. Traechtler “Anadvanced electric vehicle for development and test of
new vehicle dynamics control strategies”
Proceedings of IFAC AAC2010, Munich, July 2010 inCD9) N. Bajcinca and Y. Kouhi “Distributed optimization forfeed-forward global chassis control” Proceedings of
IFAC AAC2010, Munich, July 2010 in CD10) M. Abe “Evaluation of active vehicle motion controls
from tire energy dissipation points of view” Keynote
presentation of AVEC2012, Proceedings of 11th
International Symposium on Advanced Vehicle
Control Seoul, Korea, August 2012 in CD-10-*02-寄稿文E_*02/寄稿文 14/07/31 10:16 ページ 10

Trends of Intelligent Vehicle Dynamics Controls and Their Future11) M. Abe et al. “Tire Force Distribution Control toReduce Energy Dissipation Due to Tire Slip During
Vehicle Motion for Full Drive-by-Wire Electric”
Vehicle Proceedings of Chassis tech – plus, Munich
Germany, June 2013 in CD12) M. Abe et al. “A study on active vehicle chassis
control reducing dissipation energy by tire slip for fulldrive-by-wire electric vehicle” to be presented at 23rd
IAVSD symposium, Qingdao, China, August 201313) Jun ISHIO et al. “Vehicle Handling Quality Evaluation
through Model Based Driver Steering Behavior”
Proceeding of 20th IAVSD Symposium, Berkeley
August 2007
14) Yasuhiro Aoki, Yoshio Kano and Masato Abe“Variable Stability Vehicle with Response Parameters
Controlled by Active Chassis Control Devices”
Proceedings of 22nd IAVSD symposium, August
2011, in CD15) M. Hibi et al. “A study on steering reactive torque forsteer-by-wire vehicle using driving simulator”
Proceedings of FISITA2012, F2012-G04-006,
Beijing, China, November 2012, in CD〈 Author biography 〉

Masato Abe
Doctor of Engineering, Professor, Department of Vehicle System Engineering, Kanagawa Institute of Technology
1971
1980
1987
1995-1996
2006

Completed doctor course, School of Engineering, The University of Tokyo, Doctor of Engineering
Assistant Professor, Department of Mechanical Engineering, Kanagawa Institute of Technology
Professor, Department of Mechanical Systems Engineering, Kanagawa Institute of Technology
Visiting Professor, Mechanical Engineering, University of Leeds
Professor, Department of Vehicle System Engineering, Kanagawa Institute of Technology【Research fields】

Vehicle motion dynamics, Active control of vehicle motion, Man-vehicle system, Driving assistance, Preventive safety【Academic society and committee affiliations】
1992-Present

Editorial Board Member of VSD Journal - Official Organ of International Association for Vehicle System Dynamics

1995-2005

Trustee of International Association for Vehicle System Dynamics (affiliated to IUTAM)

1995-Present

Editorial Board Member of Proceedings for I.Mech.E. Part D

1996-2002

Chairman, Technical Paper Editing Committee, Society of Automotive Engineers of Japan

1998-2004

Trustee, Technology, General Affairs, Society of Automotive Engineers of Japan

2001-2006

Co-editor of VSD Journal

2006-2011

Vice president of International Association for Vehicle System Dynamics and assumed several positions at theSociety of Automotive Engineers of Japan including Trustee, Representative, Member/Chairman of Committees
and General Chairman of International Conferences. Fellow, Society of Automotive Engineers of Japan-11-

*03_展望E_*03_展望 14/07/31 9:53 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Perspective ]NTN Module Technology Contributes to Energy Efficiency
and CO2 Reduction in Automobiles

Takehiko UMEMOTO

In recent years the pursuit of environmental performance, throughfuel efficiency improvements for CO2 reduction and the re-use ofenergy, together with the pursuit of advanced electronic controls forsafety and comfort, is progressing rapidly in the automotive sector.
NTN are traditionally developing and supplying mechanicalcomponents such as bearings that reduce heat generation, frictionand energy loss. More recently NTN is supplying automotive
manufacturers with Energy Efficiency Improved Module productsthat comprise of several different component parts.
This article introduces NTN Module Product lines which are a significant contributor to theimprovements seen in automotive energy efficiencies.

with the peripheral components and have been
supplying them to automobile manufacturers. These
module products are contributing to lower-fueleconomy and the improvement of environmental
performance by reuse of energy and pursuit of safety
and comfort by state-of-the-art electric control.
In the following sections, I would like to introduce
our module products and their technology used in thehub, brake, AMT (automatic manual transmission),engine variable valve actuation mechanism, CVT(continuous variable transmission), and engine chain
system.

1. Introduction
Recently, fuel economy for domestic automobiles
has made big strides achieving up to 36.4km/L at JC08
mode (reviewed by the Ministry of Land, Infrastructure,
Transport and Tourism) with commercially available
automobiles in the market as of September, 2013. The
low fuel consumption is a result of implementing
several technological advancements including lightweight body, hybrid system, idling stop function, CVT(continuous variable transmission), variable valve
timing system, cylinder rest engine, etc.
In an automobile more than 100 bearings are usedthat contribute to the reduction of energy loss andheat generation, which are generated by the friction ofmany components. At NTN we are developing thinnerand lighter bearings with higher load capacity andlonger service life by adopting advanced analyticaltechnology for high efficiency, high functionality,
extremely clean materials, special heat treatment, etc.
Consequently, we announced and introduced
ULTAGE Series *1 in the market as high-function,next-generation bearings.
In addition to these high-function bearings, wedeveloped module products where the ball screw,
constant-velocity joint, gear, motor, etc are integrated

2. NTN module technology
Table 1 shows examples of our module productlines for automobiles which are currently in production
or which development is completed.
The module products are expanding to every part of
the automobile including the engine, brake, hub, andtransmission, contributing to lighter-weight, morecompactness, and higher efficiency by combining
parts or in the combining of parts with peripheral
components.
The uses and features of the module products inTable 1 are introduced next.

*1 ULTAGE is a name created from combining the words "ultimate", signifying refinement, and "stage", signifying NTN’s
intention that this series of products will be employed in diverse applications and is the general name for NTN’s new
generation of bearings that are noted for their industry-leading performance.

* Executive Officer, Automotive Business HQ/EV Module Business HQ-12-*03_展望E_*03_展望 14/07/31 9:53 ページ 2

NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles

Table 1 NTN Module product lines
Area ofapplication

Moduleproduct
nameCharacteristics

Drivetrain

Brake

Press connect spline
hub joint

Ball screw drive module
for electro hydraulic brake

Electromechanical braking unit
with parking brake

Compact
Light weight

Highly efficient regeneration
mechanism
Small form factor/light weight

Reduction of environmental load
Highly efficient regeneration mechanism
Small form factor/light weight

Appearance

Area ofapplication

Moduleproduct
nameManual transmission

EngineBall screw actuator for gear shift unit

Ball screw actuator for variable valve event& lift mechanism

Improved fuel economy
Improved comfort
High response

Improved fuel economy
Cleaner gas emission
Enhanced engine output

Characteristics
Shift side

Appearance
Select side

Area ofapplication

Moduleproduct
nameCharacteristics

Continuous variable transmission (CVT)Engine chain system

Hollow ball screw unit for driving pulley

Low friction chain lever/chain guideHigh efficiency
Light weight/compact
High response

Reduced friction loss
Improved fuel economy
Appearance

Lever
Appearance

Guide

-13-

*03_展望E_*03_展望 14/07/31 9:53 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

machined within the inner diameter of H/B, as shownin Fig. 2 (b). They are forcibly engaged by bolt
tightening and the splines within the inner diameter of
H/B are elastically deformed.
NTN calls this joint method the "Press Connect"
method. In the press connect method, the engaged
surfaces are completely mated and transmit thetorque with the entire surface of the teeth, therefore,
the length of the spline engagement can be reduced.
Since this method uses elastic deformation forengagement, the same condition can be obtained nomatter how many times they are engaged. By taking
advantage of this feature, we were able to downsize
the components by methods such as thinning theinner diameter part of the hub ring and bolt fastening
mechanism to achieve a lighter underbody (approx.
12% per vehicle). This technology was only madepossible by turning the H/B and CVJ into modules withhighly efficient fastening.

3. Press Connect Spline Hub Joint
PCS-H/J
The power from the engine is transmitted to the hub
bearings and tires through the drive shaft, as shown in
Fig. 1. Currently, the hub bearings (H/B) and constantvelocity joint (CVJ) are engaged through splines
machined on both CVJ stem and inside H/B.
The torque is transmitted through the spline teeth
surface. This surface is not completely mated becauseof the difference of tooth pitch due to the processaccuracy. This gap between the teeth leads tobacklash and the length of engaged splines must be
increased in order to obtain the required torquecapacity.
With the new technology of press connect spline
hub joints (PCS-H/J), a special spline (groove andraised feature) is machined on the CVJ stem and a
spline with tooth width smaller than the CVJ stem is

Drive shaft: constant-velocity joint (CVT)
Hub bearing (H/B)
Tire

Fig.1 Drivetrain(b) Press connect spline hub joint
PCS-H/J(a) Current structure
Third-generation hub bearing (H/B) + constant-velocity joint (CVJ)
100
45

100

φ80

φ80

45

Fastenedwith a nut
Fastenedwith a bolt

100

Shorter CVT stem

35
H/B

Spline boundary

W1
W2

Gap

H/B

W2
W1

W1 ＞ W2

W1 ＜ W2

No gap on the tooth surface

Gap on the tooth surface

Mass: 3.88 kg (reduction: 0.94 kg/unit)

Mass*2 : 4.35kg*2: Mass of H/B + outboard CVT outer ring + nut or bolt

Fig.2 Comparison with “Gen 3H/B+CVJ” and PCS-H/J-14-

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NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles

Driver

4. Ball Screw Drive Module for Electro
Hydraulic Brake

Braking force

The regenerative brake was adopted for electric andhybrid vehicles to recover kinetic energy as electricenergy while the brake is applied. In order to recover
the most possible electric energy, it is necessary tocreate a mechanism that uses the motor braking force(regenerative brake) to the maximum and supplement
it with hydraulic braking force.
The cooperative control between the regenerative
brake and hydraulic brake is also critical, so that thesystem can immediately respond to the changing
braking force.
Fig. 4 shows the construction of Electro HydraulicBrake (EHB) with a regeneration cooperative brakemechanism. Our “ball screw drive module for electrohydraulic brake” (Fig. 5) was adopted in massproduction vehicles.
The mechanism for this module is explained using
the system configuration in Fig. 4 and as follows:
When a driver presses the brake, the pedal stroke
signal is sent to ECU and the optimum braking force
and hydraulic force are calculated. ECU operates theball screw drive module for the electro hydraulic brakewith a DC motor and moves the piston in the hydrauliccylinder to the optimum extent. The hydraulic force
produced from the hydraulic cylinder is transmitted tothe caliper of each wheel to apply the braking power.
By using the ball screw drive module, quick andprecise control and the maximum energy recovery canbe achieved. This significantly increased the fueleconomy compared to conventional systems
achieving the highest fuel efficiency level in the world.

ECU

DC motor

Hydraulic braking force Regeneration braking force

Ball screw
drive module

InverterCaliper

Hydraulic cylinder

Caliper

Motor

Fig.4 Constitution of electric hydraulic brakeIN

Aluminum caseDC motor rotational torque
Ball screw

Bolt
Idler gear
Output gear
SleeveOUT
Thrust power(hydraulic cylinder is pushed)

Support bearings

Fig.5 Cross section structure of ball screw drive module

Driver's required braking force
Braking force by hydraulic brake

Braking force by motor(regenerative brake)

Start of braking

Force applied bybrake pedal
operation

End

Fig.3 Regeneration cooperative brake-15-

*03_展望E_*03_展望 14/07/31 9:53 ページ 5

NTN TECHNICAL REVIEW No.81（2013）is proposing its adoption.
The mechanism of electromechanical brake andparking brake is explained using the schematic ofelectromechanical braking unit in Fig. 6. The DC
motor attached to the unit drives the rotation of thesun roller in the electromechanical braking unit
through the gear reducer. This driving force to the sunroller drives autorotation and revolution of the adjacentplanetary roller moving the screw between theplanetary roller and the outer ring which drives theouter ring itself toward the axial direction. This axial
force of the outer ring is transmitted to the brake pad
producing the braking force.
The parking brake mechanism is equipped
immediately below the intermediate gear. The pin
inside the solenoid is inserted to the ratchet groove
set on the reverse side of the intermediate gear toactivate the parking brake and is pulled out to release
the brake.

5. Electromechanical Braking Unit with
Parking Brake
The regenerative brake technology which recovers
kinetic energy as electric energy when the brake isapplied is expanding mainly with hybrid and electriccars for improved fuel economy.
Commercially available automobiles today are
equipped with hydraulic brake systems; however, as amethod for the efficient recovery of energy, EHB(introduced in the previous section) would contributeto highly efficient regeneration.
There is another technology for improving thecurrent hydraulic brake system for highly efficientenergy recovery by directly operating the friction brakecalled Electromechanical Brake (EMB).
Having already developed an electromechanicalbraking system, NTN developed “Electromechanical
Braking Unit with Parking Brake” by adding a parking
brake mechanism to the electromechanical brake andOuter ringLinear motion
mechanism

Caliper

X

Y

107m

m

Z

Gears

Sun roller
62mm

Brake pad

Planetary roller

Rotation

ThrustMotor
84.5mmBrake disk

Section Y-YXYSolenoidZ
Section X-X

MotorIntermediategear
ONSection Z-Z(parking brake mechanism)

OFF

Fig.6 Schematic of EMB unit

-16-

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NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles

The mechanism of operation is explained using thestructure diagram of the gear shift unit in Fig. 9. Theball screw actuator on the select side inside the unit
moves the striker to the axial direction and then theball screw actuator on the shift side oscillates thestriker to shift the gear to the selected speed. The
gear shift is made by cooperative control of theorthogonal ball screw actuator with the motor. The
response is very fast with an operation time of 0.06
sec in the shift side and 0.08 sec in the select side.
This AMT is characterized by a compact size
without significant change in the structure by installing
the gear shift unit on top of the regular manualtransmission.

6. Ball Screw Actuator for AutomatedManual Transmission Gear Shift Unit
Vehicles with automated manual transmission(AMT) are expanding for improved fuel economy andcomfort. There are different designs for the structureof AMT. Here, we are describing AMT with a gear shift
unit which adopts NTN-developed ball screw
actuators.
Since 2004 this gear shift unit has been adopted for
mid to small size trucks with a frequent start/stop
trend. Fig. 7 shows an external view of AMT and Fig.
8 shows how the gear shift unit is installed.
Fig. 9 shows the structure of the gear shift unit andFig. 10 shows an external view of the ball screw for
the actuator.

Shift side

Select side

Fig.7 Automated manual transmission

Shift sensor

Link arm

IN
Ball screw on shift side

Fig.8 Gear shift unit

Select sensor

Shift side

Striker

Select side
Ball screw on select side

Fig.10 Actuator for ball screw

Ball screw actuator

Fig.9 Gear shift unit

-17-

*03_展望E_*03_展望 14/07/31 9:53 ページ 7

NTN TECHNICAL REVIEW No.81（2013）This mechanism makes instantaneous and precise
control possible due to the adoption of the ball screw
actuator and achieves improved engine output andresponse, improved fuel economy of about 10%, andcleaner gas emission.
Fig. 11 shows the variable valve event & lift (VEL)
mechanism and Fig. 12 shows the structure of the ballscrew actuator. The translatory movement of the nutof the ball screw actuator directly connected to the DC
motor controls the rotational phase of the control shaft
which is converted to vertical movement of the valve
lifter.
This mechanism variably and continuously controls
the valve lift and valve event by precisely controllingthe movement of the ball screw actuator nut with theposition sensor.

7. Ball Screw Actuator for Engine Variable
Valve Event & Lift Mechanism
For global environmental improvement andenhanced power output, technological development isin progress for refined engine control in order toachieve the optimum response to driving conditions.
One of the technologies for the optimum control is thevariable valve event & lift (VEL) mechanism. Improved
fuel economy and power output are pursued by
optimizing the valve open/close timing and the lift
volume at the top of the engine. The engine variable
valve event & lift mechanism, which adopts our ballscrew actuators, has been mass produced anddeployed in the market since 2007 for passenger
vehicles, sports cars, and RVs.

Actuator
Position sensor
DC motor

Ball screw nut

Ball screw shaft

Rocker arm
Link A
Control shaft

Support bearing
Link B

Output cam

Eccentric shaft
Drive shaft
Valve lifterBall screw
Valve

Fig.11 Appearance of VEL system

Yoke
Control shaft

Coupling
DC motor

Nut
Housing

4-point contact ball bearing

Ball screw shaft

Fig.12 Structure of ball screw actuator-18-

Deep groove ball bearing

*03_展望E_*03_展望 14/07/31 9:53 ページ 8

NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles

can be installed coaxially with the pulley driving
mechanism. This is a world's first technology that
achieved ball circulation at the screw shaft side for theball screw circulation mechanism.
The example of the structure in Fig. 14 describes amechanism where the rotational power from the DC
motor is transmitted to the hollow screw shaft and thenut fixed to the housing moves the screw shaft with adriving pulley toward the axial direction for changing
speed.
By the adoption of the “Hollow Ball Screw Unit," themotion force and electric power loss associated withthe pulley drive and the DC motor size can be reduced
as well as the downsizing of the entire electroniccontrolled CVT driving mechanism.

8. Hollow Ball Screw Unit for Continuous
Variable Transmission (CVT)The continuous variable transmission (CVT) *3,
which contributes to fuel economy as well as smooth
speed shifting, is currently adopted by many vehicle
models.
The pulley driving method of the CVT transmission
mechanism is now changed to electronic control fromthe conventional hydraulic control for more precise
speed control and reduced power loss. However,since the current improved electronic control still uses
the sliding screw for pulley drive, there is further room
for improvement in terms of power loss.
The "Hollow Ball Screw Unit” developed by NTN(Fig. 13) converts the rotational power of the DC
motor into axial movement of the pulley with a highefficiency of 70% or more. Due to the light weight,
thin, and compact construction of the hollow axis, it*3 CVT: An abbreviation of Continuously Variable Transmission.

The system for changing the speed by adjusting thepulley width by the pulley driving mechanism.

DC motor

Ball
Nut

Gear 1

Ball circulation
grooveGear 2

Hollow
screw shaft

Driving pulley

Screw shaft

Fixed pulley
Belt

Gear 3

Fig.14 CVT mechanism
Ball circulation mechanism
Fig.13 Hollow ball screw unit

-19-

Nut

*03_展望E_*03_展望 14/07/31 9:53 ページ 9

NTN TECHNICAL REVIEW No.81（2013）

loss indicates 27% at the chain lever and 13% at the
chain guide which suggests a large ratio of friction
loss due to the sliding operation of the chain and thelever guide.
We have developed and proposed in the market the
low-friction chain lever/chain guide by converting thesliding of the resin lever guide into rolling as shown in
Fig. 16.
The low-friction chain lever/chain guide is anaddition of the small and highly rigid bearing roller unit
onto the current resin-made lever guide. As shown in
Fig. 17, the chain does not directly contact the lever orguide but rolls over the bearing roller unit resulting in
22% reduction of friction loss.
The adoption of this low-friction chain lever/chain
guide resulted in 1% *4 of improvement in fueleconomy for the entire vehicle.

9. Low Friction Chain Lever/Chain Guide
It is important to transmit generated power with thetiming chain, precisely drive cam shaft, and produce
exact valve timing to achieve a high performanceengine. Synchronization accuracy and durability areneeded for transmitting power and a chain system iscomposed of the “timing chain”, “sprocket”, “tensioner”,“chain lever” and ”chain guide” as shown in Fig. 15.
The friction loss in diverse parts of the engine must
be reduced as much as possible for effectively using
the generated power.
The ratio of the loss generated in different
components is 80% in the crankshaft including
pistons, 4% in the cam shaft, and 16% at the chain
systems. 1)
The breakdown of the second largest chain system*4 Calculated value based on experimental data

SprocketTiming chain

Tensioner

Low friction productLever side

Current product

Guide side

Chain lever

Chain guideCam shaft

Crank shaft

Fig.16 Chain lever and guideFig.15 Timing chain system

Lever

軸受ローラユニット
Guide

Fig.17 Low friction chain lever and guide-20-*03_展望E_*03_展望 14/07/31 9:53 ページ 10

NTN Module Technology Contributes to Energy Efficiency and CO2 Reduction in Automobiles

10. Conclusion
I have introduced NTN products already in mass
production and adopted by our customers as well as
the products which are being developed and currently
being proposed to customers among our module
technologies and products that contribute to theimprovement of energy efficiency of vehicles.
Traditionally, NTN has been contributing to the highfunctionality and high energy efficiency of individual
elements such as bearings, CVJ, and ball screws.
NTN considers that further enhancements,
improvements, and higher functionality are
increasingly important by utilizing module technology
where the peripheral components are also integrated.
In addition to the improved fuel economy, the pursuitof safety and comfort is becoming quite important in
the automotive community.
NTN is striving to continue development of
combined technology consisting of our existing
technologies such as tribology, machine technology,
and mechatronics technology for contributing not only
to the vehicles with low-fuel consumption engines but
also to the next generation vehicles such as EV, HEV,and FCV.

Reference
1) Exhibit: Society of Automotive Engineers of Japan,
Academic Conference preprint issue, No. 55-00

Photo of author

Takehiko UMEMOTO
Executive Officer
Automotive Business HQ
EV Module Business HQ-21-

*04E_03 14/07/31 10:27 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Article ]Recent Technology Trends of In-Wheel Motor System
for Automotive
Tomoaki MAKINO*
Aiko ISHIKAWA*
Chinami ITOU*
Kayo SAKAI*In the electric vehicles equipped with in-wheel motors, the degrees of
freedom for vehicle design are high because there is no restriction ofdrivetrain layouts such as drive shaft. Moreover, independent, quick
and precise control of driving force of each wheel can be realized by inwheel motor system, meaning that the driving stability can be improved significantly. This paper
describes the technology trends of in-wheel motor system expected to be powertrain for new
generation electric vehicle.

1. Introduction

2. Dawn of in-wheel motor electric vehicles

Electric vehicles have recently been drawing
attention as possible measures against environmental
issues such as global warming, air pollution, andenergy issues such as the depletion of fossil fuel.
In-wheel motor electric vehicles have driving motors
in the wheels, so the power is transmitted to the tires
without needing a differential gear and drive shaft.
There is no restriction of the layout caused by thedrive train, enabling a higher degree of freedom forthe vehicle design. In addition, the vehicle
maneuverability can be greatly increased since thedriving power can be quickly and accurately controlled
for each wheel independently.
NTN has been developing an in-wheel motor
system with high mobility functions combining
independent wheel driving and steering control as well
as a reducer mechanism that effectively contributes tolight-weight vehicles.
In this article, we introduce the technology trend of
in-wheel motors which holds a high expectation as thedriving method for the next-generation electricvehicles and includes NTN’s developments.

The history of in-wheel motor vehicles is long. In
1899, Dr. Ferdinand Porsche already developed aseries of hybrid vehicles which drove the front wheels
with a motor powered by electricity generated by the
installed engine. Fig. 1 shows the structure of the inwheel motor patented by Dr. Porsche. It used a directdrive method with no reducer and the in-wheel motorof essentially the same structure was installed in thevehicle mentioned above.

Fig. 1 Structure of the first in-wheel motor 1)* Drive System Engineering, EV Module Business HQ-22-

*04E_03 14/07/31 10:27 ページ 2

Recent Technology Trends of In-Wheel Motor System for Automotive

3.1 Direct drive system

Back in those days, performance of the battery andmotor was not high enough and had a short driving
range so the vehicle performance was insufficient.
Electric vehicles started losing their competitive edge
with the development of oil resources and improved
availability of gasoline as well as the cost reduction ofreciprocal engine vehicles starting with the mass
production of the Ford T model in 1908.
Subsequently, the era of full-scale development ofmotorization with reciprocal engine vehicles started. In
the 1970s, oil supply concerns presented by the oilcrisis and environmental issues such as air pollution
due to exhaust gas started emerging. Electric vehicles
were looked as a measure against those issues;
however, they did not evolve into one due to problems
such as battery performance, etc.
In the 1990s and thereafter, performance of electricvehicles significantly increased due to thedevelopment of (1) a lithium-ion battery which has
higher energy/output density compared to theconventional lead battery, (2) a permanent magnet
synchronous motor using neodymium magnet and (3)inverter-based motor control technology. As the global
warming issues due to greenhouse gas such as CO2
are widely recognized, electric vehicles and thedevelopment of in-wheel motor is attracting attention
in recent years.

Many of the direct drive systems developed so far
adopt the outer rotor method which places the rotatoron the outside of radial direction from the stator 2)-4).
Fig. 3 shows an example of its structure. A case ofdevelopment of vehicles with in-wheel motors of thissystem is also reported 3). Fig. 4 shows an
appearance of such a vehicle.

Rotor

Wheel

StatorBracket (2)Bracket (1)Brake disc

Axle hubFig. 3 An example of the structure of direct driving
system in-wheel motor 2)

3. Structure of in-wheel motor
As shown in Fig. 2, drive systems of electricvehicles are broadly divided into the onboard system,
which installs a motor in the body and transmits thepower to the tires through a drive shaft, etc, and an inwheel motor system which installs motors inside thewheels.
The in-wheel motor system is further subdivided intoa direct drive system, which does not equip a reducer
and a system which employs a reducer concurrently.

Fig. 4 Developed vehicle equipped withdirect driving system in-wheel motors 3)

3.2 Reducer system
Although using a reducer brings somedisadvantages such as complex structure with anincreased number of components, the torque required
for the motor can be reduced which allows for asmaller motor size resulting in a more compact andlighter in-wheel motor as a whole.
Many cases of development of in-wheel motors of
this method are reported 5)-9). Fig 5 and 6 show someexamples of those cases. The example of Fig. 5
adopts a reducer with the planetary type reduction
gear and the example of Fig. 6 adopts both the
reducer with planetary type reduction gear and theparallel shaft type reduction gear.

Electric vehicle
In-wheel motor system

Reducer system

Onboard system

Direct drive system

Fig. 2 Classification of electric vehicle by driving system-23-

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NTN TECHNICAL REVIEW No.81（2013）

Fig. 7 and Fig. 8 show the basic structure of cycloid
reducer, which is K-H-V type planetary type reduction
gear, and the structure of in-wheel motor that NTN
developed using this reduction gear 10), 11)
respectively. This reduction gear consists of acycloidal disc that has an epitrochoid curve andmultiple ring gear pins placed at equal intervals on thecircumference. Only the rotation movement of theeccentrically revolving cycloidal disc is transmitted tothe output shaft, located inside the cycloidal disc.
Since the cycloid reducer is able to provide a large
reduction ratio with only one stage, it is superior toother reducer systems for its compact size and lightweight.

Motor

Ring pin
Cycloidal disc

Brake

Hub

Gear

Input shaft

Output shaft

Fig. 7 Cycloid reducer
Wheel

Fig. 5 An example of the structure of reducer system
in-wheel motor 8)

Reducer housing

Cycloid reducer

Cooling fin

Motor housing

Wheel
Motor stator

Brake disc

Motor rotor

ResolverCounter gear

SplinePlanetary gear

Hub

Hub bearings

Outputshaft
Reducer
input shaft

Pump
Lower armball joint

Hub part

Oil pump

Reducer part

Stator

Rotor

Rotor shaft

Motor partFig. 8 In-wheel motor with cycloid reducer 11)
Fig. 6 An example of the structure of reducer system
in-wheel motor 9)-24-

*04E_03 14/07/31 10:27 ページ 4

Recent Technology Trends of In-Wheel Motor System for Automotive

4. Impact of unsprung mass

5. Vehicle motion control

In general, an in-wheel motor system increases theunsprung mass compared to an onboard system,
which adversely affects the comfort and road-hugging
property of the tires.
The basic oscillation model for a suspension system
includes a two-degree of freedom oscillation model
considering sprung and unsprung masses. Fig. 9shows the sprung mass acceleration based on thisoscillation model. As the unsprung mass increases,
the unsprung resonance frequency decreases and the
sprung acceleration in the low frequency range
increases 12). Humans are most sensitive to a verticalvibration of 4-8 Hz 13). Therefore, an increase of the
sprung acceleration in this range due to increase of
unsprung mass significantly reduces the comfort level.
It also reduces the followability of unsprung mass tothe irregularities of the road surface which adversely
affects the road-hugging property of the tires.
Several reports on the impact of in-wheel motors to
the increase of unsprung mass are available 14), 15) so
reducing the weight of in-wheel motors to reduce its
impact is critical. In addition, an initiative to address
this unsprung mass issue by a mechanism is also
underway. For example, a mechanism was developed
to prevent adverse effect of riding comfort configuring
the motor mass with a movable dynamic damper thatis mounted vertically 16).
NTN achieved a significant reduction of in-wheel
motor weight by integrating the cycloid reducer as
mentioned above.

The drive motion control by an electric motor has
the following advantages over internal combustion
engines.
(1) High torque response(2) Highly accurate torque control(3) Ability to generate and control braking force(4) Independent control of driving force for eachwheel by mounting motors in them
In addition, an in-wheel motor system allows further
high-response torque control since it is not affected byresonance of the power train such as the drive shaft 12).
It is also able to control braking force with an evenhigher response than ABS (anti-lock brake system)and hydraulic brakes for traction control used withinternal combustion engine vehicles 17).
By using the above properties, particularly electricvehicles with an in-wheel motor system, can control
highly advanced vehicle motion which is difficult withinternal combustion engine vehicles.

5.1 Anti-slip control
Fig. 10 shows the characteristics of lateral andlongitudinal (driving or braking) forces between thetires and road. The characteristics of Fig. 10 can betheoretically derived from a model to consider the tires
as infinite elastic bodies independently placed on the

4.0

2.4

Increase of unsprung mass

5˚

1.6

Unsprung resonant point

30˚
20˚

10˚ 3˚

0.8

Sprung resonant point

1˚

Comfort zone

-4.0

Longitudinal force/lateral force kN

Sprung acceleration

Side slip angle

Lateralforce
kN

BenchmarkDiscomfort zone
Resonance frequency decreasedby the increase of unsprung mass
0

Vertical load 3.8 kN

10

20

Excitation frequency Hz

Fig. 9 Effect of unsprung mass on vehicle sprung
acceleration 12)

-3.2

-2.4

-1.6

-0.8

Braking force kN

4.0
3.2

1.6

3˚

2.4

3.2

Driving force kN

4.0

Vertical load 3.8 kN

1˚

5˚

Longitudinal force

10˚

1.6

0

0.8

(a)

20˚

2.4

0.8

0

1˚

3˚

5˚

0.2

10˚

20˚

0.4

30˚

30˚

Side slip angle

0.6

Longitudinal slip ratio

Lateral force

0.8

1.0

(b)
Fig. 10 Tire characteristics with driving and braking 18)-25-

*04E_03 14/07/31 10:27 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

circumference. The longitudinal force produced at thetires increases along with the slip ratio and decreases
after reaching the peak. Conversely, the lateral force
decreases as the slip ratio increases 18).
As the road condition changes and an excessive
slip occurs on the tires, both the driving force andlateral force decrease and may cause not only the
reduction of acceleration performance but also theloss of directional stability. Therefore, prevention ofexcessive slip of tires is critical in order to avoid
unstable vehicle behavior.
By using the motor driving properties mentioned
earlier, electric vehicles can achieve a more superioranti-slip control than internal combustion engine
vehicles. Particularly, an in-wheel motor system canprovide the most stable control. Several researches
were made and reported regarding anti-slip control
using an in-wheel motor system, such as the feedback
control using the wheel speed and slip ratio control
aiming the maximization of driving/braking force 19)-21).

Fig. 11 Test vehicle with four wheel independent driving
and steering

Table 1 Main specifications of test vehicle
Body size

2600×1560×1350mm (Length×width×height)

Mass

619kg

Drive system

Four-wheel drive (in-wheel motor)

Steering system Four-wheel independent steering (steer-by-wire)

5.2 Vehicle stability control
Under the condition where the tire slip ratio isrelatively small and the degree-of-freedom for
driving/braking force control is relatively high, thevehicle driving stability can be increased byindependently controlling the driving/braking forces ofeach wheel. For example, since yaw moment can bedirectly controlled by giving different driving forces tothe right and left wheels, the vehicle maneuverability
such as the yaw rate followability during steering
increases 22), 23).
In addition, the traveling performance can beincreased by appropriately controlling the driving forceof each wheel. For example, turning performance canbe increased by allocating a larger driving force to thewheel which has allowance for tire grip 24), 25). This
type of performance improvement can be further
increased by controlling the driving force and steering
altogether.
Fig. 11 and Table 1 shows the appearance andmain specifications of a test vehicle NTN created. This
vehicle has in-wheel motors on all four wheels and thecapability to independently steer these four wheels.
NTN and Kanagawa Institute of Technology identified
that by independently controlling the driving force andthe steering angle the energy dissipation of the tires isreduced 26), the tire load rate is decreased, and thevehicle limit performance is improved 27).
In addition, with an in-wheel motor system thevertical load produced by the driving force through thesuspension mechanism becomes larger than a vehicle
with a drive shaft since the knuckle that supports thein-wheel motor receives the torque reaction 12). By

taking advantage of this property, the vertical body
motion due to the driving power action can be more
effectively and independently controlled for eachwheel than with drive shaft-driven vehicles, thus
improving response to the roll and yaw motions 28).
The comfort can be also improved by reducing thevehicle roll and pitch motions 24).

6. Next-generation mobility
The in-wheel motor system does not require a
power train such as a drive shaft for transmitting
power. Therefore, vehicle structure and motion noteasily accomplished by conventional vehicles can now
be achieved.
Fig. 12 shows an example of in-wheel motor
application for a two-passenger commuter vehicle
designed for short/mid range mobility 29). By installing
drive motors within the wheels, plenty of passenger
and cargo space can be secured despite the size of its
compact body. In February 2013, the Ministry of Land,
Infrastructure, Transport and Tourism announced thestart of a certification program for a "ultra-small
mobility” vehicle standard for two-passenger cars
smaller than a “light motor vehicle.” This is a trial
measure for establishing a new vehicle standard
expected to be applied for elderly and child-support,
tourism and regional promotion, etc. With theestablishment of the dedicated standard for thiscategory of vehicles, it is expected that their
expansion will be accelerated in the future.
-26-

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Recent Technology Trends of In-Wheel Motor System for Automotive

In addition, by combining an in-wheel motor andindependent wheel steering mechanism, “sideway
motion” and “pivot turning” can be achieved. Fig. 13
shows an example of such a vehicle and Fig. 14
shows the steering for each traveling mode. For “pivot

turning," the wheels are steered so that the axis ofrotation of each wheel points to the center of thevehicle. For "sideway motion,” all of the wheels aresteered perpendicular to the ordinary forward
direction 30).

(a) Without cowl 29)

Fig. 13 In-wheel motor type electric mobility(b) Registered vehicle with cowl
Fig. 12 In-wheel motor type two-seat commuter

Front left wheel(a) Normal forward(b) Pivot turn：Center of rotation

Fig. 14 Steering status at each running mode 30)-27-(c) Sideway motion

*04E_03 14/07/31 10:27 ページ 7

NTN TECHNICAL REVIEW No.81（2013）10) MINORU SUZUKI, et al.: Development of In-Wheel
Motor Type Axle Unit, NTN TECHNICAL REVIEW, 75(2007) 46.

7. Conclusion
In this article, we have described the structure of
typical in-wheel motors, initiative for reducing impactof unsprung mass increase and the technologicaltrend of vehicle motion control using in-wheel motors
while focusing on the recent development trend. Inwheel motors can realize a level of vehicle motion
which is difficult to achieve with conventional vehicles,
which makes it an ideal drive system for the next
generation electric vehicles. NTN will further improve
both the mechanism and control aspects of thesystem with an aim for early commercialization.

11) YUICHI ITOH, KAYO SAKAI, YUSUKE MAKINO: InWheel Motor System, NTN TECHNICAL REVIEW, 79(2011) 22．
12) HIROAKI YOSHIDA: Evolution of In-Wheel Motor
System, Journal of Society of Automotive Engineersof Japan, Vol. 65, No. 3 (2011) 80.
13) Society of Automotive Engineers of Japan:
Automotive Technology Handbook 1,
Introduction/Theory (1990) 269.
14) M. ANDERSON et al.：Unsprung Mass with InWheel Motors – Myths and Realities，Proc. AVEC’10，(2010) 261．
15) R. ROEL et al.：Influence of in-wheel motors on theride comfort of electric vehicles，Proc. AVEC’10，(2010) 835．

References
1) F. PORSCHE：Improvements in or relating toElectrically Driven Road Vehicles, G. B. Patent 18099(1900)．

16) GO NAGAYA, YASUMICHI WAKAO, AKIHIKO ABE:
Development of an –In-Wheel Motor with Advanced
Dynamic-Damper Mechanism, Proceedings of theSociety of Automotive Engineers of Japan, No. 83-02(2002) 9.

2) YOSHIHIRO OKAMOTO, TOMOAKI KIRITANI,YUKIHIRO NODA: Development of High Efficiency Inwheel Motor System for using 20-inch Wheel Mounted
Motor, Toyo Denki Review, Vol. 117 (2008), 12.

17) F. BOTTIGLIONE, A. SORNIOTTI & L. SHEAD：The
Effect of half-shaft torsion dynamics on theperformance of a traction control system for electricvehicles，Proc IMechE PartD : J Automobile
Engineering, 226, 9 (2012) 1145．

3) TOMOYUKI SHINKAI, et al.: Development of an
Electric Vehicle “SIM-WIL” Journal of Society ofAutomotive Engineers of Japan, Vol. 66, No. 9 (2012)84.

18) MASATO ABE: Automotive Vehicle Dynamics (2nd
Edition), Tokyo Denki University Press (2012) 38.

4) A. WATTS et al.：Integrating In-Wheel Motors into
Vehicles-Real-World Experiences，SAE Int. J. Alt.
PowerI, 5.1 (2012) 289．

19) YOICHI HORI: Future Vehicle driven by Electricity
and Control -research on four-wheel-motored,
SEISAN KENKYU Vol. 56, 2 (2004) 131.

5) SHIRO MATSUGAURA, KIYOMOTO KAWAKAMI,HIROSHI SHIMIZU: A Proposal on Chassis Structure
and Performance Evaluation for an Electric Vehicle,
Transaction of Society of Automotive Engineers ofJapan, Vol. 35, No. 3 (2004) 117.

20) MOTOKI SHINO, YOSHIAKI OHSHIMA, MASAONAGAI: Wheel Velocity Control of Micro-Scale
Electric Vehicle for Improving Directional Stability,
Transactions of the Japan Society of MechanicalEngineers Series C, Vol. 70, No. 694 (2004) 1680.

6) YASUAKI TAHARA, et al.: Development of the InWheel Motor unit, Proceedings of the Society ofAutomotive Engineers of Japan, No. 131-06 (2006) 9.

21) HIROSHI FUJIMOTO, MASATAKA YOSHIMURA:
Driving Force Control for Electric Vehicle Based on
Slip Ratio Control, Proceedings of the Society ofAutomotive Engineers of Japan, No. 3-11 (2011) 11.

7) R. S. ZHOU & F. HASHIMOTO：Highly Compact
Electric Drive for Automotive Applications，SAE
Technical Paper 2004-01-3037．

22) MAKOTO KAMACHI, KEVIN WALTERS, HIROAKIYOSHIDA: Improvement of Vehicle Dynamic
Performance by Means of In-Wheel Electric Motors,
Mitsubishi Motors Technical Review, No.18 (2006)
107.

8) YUTARO KANEKO, et al. : Development of compact
and high power In-Wheel-Motor unit, Proceedings of
the Society of Automotive Engineers of Japan, No. 709 (2009) 11.
9) SATOSHI MURATA: Development of In-wheel-motor
drive unit, Proceedings of the Society of Automotive
Engineers of Japan, No. 28-10 (2010) 5.

23）F. BRAGHIN & E. SABBIONI：Development of acontrol strategy for improving vehicle safety in a
hybrid vehicle with four independently driven in-wheel
motors, Proc. AVEC’10, (2010) 91．

-28-

*04E_03 14/07/31 10:27 ページ 8

Recent Technology Trends of In-Wheel Motor System for Automotive24) DAISUKE AKAHO et al.: Development of Vehicle
Dynamics Control System for In-Wheel-Motor
Vehicle, Proceedings of the Society of Automotive
Engineers of Japan, No. 120-10 (2010) 1.
25) J. KANG, Y. KYONGSU & H. HEO：Control
Allocation based Optimal Torque Vectoring for 4WD
Electric Vehicles，SAE Technical Paper 2012-010246．
26) RYOJI NAKAJIMA et al.: The Effects of tire force
distribution control on reducing tire slip energy
dissipation, Proceedings of the Society of Automotive
Engineers of Japan, No. 30-13 (2013) 21.
27) YUDAI SUZUKI et al.: A study on effects of tire force
distribution control on vehicle motion with Full-DriveBy-Wire EV, Proceedings of the Society ofAutomotive Engineers of Japan, No. 30-13 (2013) 25.
28) ETSUO KATSUYAMA: Decoupled 3D Moment
Control by In-Wheel Motor, Proceedings of theSociety of Automotive Engineers of Japan, No. 3-11
(2011) 1.
29) TETSUYA YAMAMOTO, AIKO ISHIKAWA, WATARU
YAMADA: Development of In-Wheel Motor System
for Electric Commuters, NTN TECHNICAL REVIEW,
No. 79 (2011) 29.
30) AKIRA YAMAGATA et al.: Development of Two-seatElectric Mobility with In-Wheel Motors NTN
TECHNICAL REVIEW, No. 79 (2011) 51.

Photo of authors (titles are at the time of development)

Tomoaki MAKINO

Aiko ISHIKAWA

Chinami ITOU

Kayo SAKAI

Drive System Engineering
EV Module Business HQ

Drive System Engineering
EV Module Business HQ

Drive System Engineering
EV Module Business HQ

Drive System Engineering
EV Module Business HQ-29-

*05E_03 14/07/31 13:48 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Electro-mechanical Brake Unit with Parking Brake

Tatsuya YAMASAKI*
Makoto MURAMATSU*
Yui MASUDA**The new brake technology such as regenerative braking and Electric
Parking Brake (EPB) has been spreading in order to meet the demands toimprove the fuel economy and convenience of vehicles. It needs to utilize
Electro- mechanical Brake (EMB) which is the final form for the further
advanced features of brake.
NTN has developed the EMB unit which has parking brake function.

use coordinated control technology with a service
brake which uses friction for obtaining braking force.
The commercial vehicles control the service brakeusing an electric-powered hydraulic actuator 2)-5),
however, not only brake oil that has highenvironmental load is used with the current hydraulicsystem. Due to the insufficient response, it is alsodifficult to maximize recovery of kinetic energy. As ameans to solve this issue, brake manufacturers aredeveloping an electro-mechanical brake (EMB) that
operates the service brake by mechanically controllingbrake pistons using rotation of the motor.
NTN has been developing and is proposing compact
and light actuators for electro-mechanical brakes using
our proprietary linear motion mechanism 6)-9). We have
now developed an electro-mechanical brake unit witha newly added parking brake function and revisitedthe internal design to further reduce its compact sizeand enhance its performance. In this article, we are
introducing the structure of the improved actuator andcharacteristics of the electro-mechanical brake unit.

1. Introduction
Electric parking brakes (EPB) 1) that operate theparking brake by an electric actuator and aregenerative brake which recovers vehicle kinetic
energy as electric energy by motor, etc. are growing
among brake related technologies.
Two types of EPBs are commercially available,
namely, the type that installs a dedicated electric
actuator on hydraulic brake caliper and the type thatpulls the wire of the conventional parking brakemechanism with an electric actuator. EPBs haveseveral advantages such as improved convenience
with automatic uphill start assistance and spacesaving in the cabin by replacing the parking brakelever with a switch.
A regenerative brake is a technology to recover thevehicle kinetic energy as electric energy when thebrake is applied. This helps to improve the fuelefficiency (electric cost) and its use has been
expanding mainly with the hybrid and electric vehicles
which use motors for driving power. In addition, somevehicle engines which recover the vehicle kinetic
energy by alternators, etc. and use the energy as thepower source for auxiliary equipment are also recently
being commercialized. In order to efficiently recoverenergy using a regenerative brake, it is important to*Chassis System Engineering, EV Module Business HQ**Control Systems Engineering, EV Module Business HQ-30-

*05E_03 14/07/31 13:48 ページ 2

Electro-mechanical Brake Unit with Parking Brake

invention is used for the linear motion mechanism.
This mechanism has a high load conversion ratio(ratio of output pressure over the input torque);therefore, the reduction ratio of the gear train can be
low which contributes to the small size of the entire
unit.
In this development, we have revisited eachelement of design to reduce the volume of theactuator by 30% compared to the conventional
product (developed in 2011) and in addition, improved
the response, as described later.
Furthermore, we have added a parking brakemechanism to the gear area as a new function. With
this function, no separate parking brake is necessary
for the vehicle with expected improvement on theassembly work and cost. The parking brakemechanism is described in the next section in detail.

2. Structure of electro-mechanicalbrake unit with parking brake
2.1 Electro-mechanical brake unit
Fig. 1 shows an appearance of an electromechanical brake unit developed for the rear wheelsof 1500cc-class vehicles and Fig. 2 shows its internal
structure. In addition, Table 1 and 2 show thespecifications of the unit and the motor used in theunit.
This unit has a linear motion mechanism and a
motor equipped in parallel and parallel shaft gears areused for the transmission of torque between them.
The torque produced by the motor is transmitted to thelinear motion mechanism via the gears and converted
to the load (hereinafter, “pressure”) to press the pad
onto the disc. To reduce the size of the entire unit, aplanetary roller screw mechanism 6) of NTN's own

Caliper

Linear motion
mechanism

Top view

Motor
Gear part

Back view

Isometric view

Side viewFig. 1 EMB unit

-31-

*05E_03 14/08/01 16:19 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

Caliper Linear motion
mechanism

Brake pad

Y

X

m

Motor
84.5mmBrake disc

Input gear

62mm

Output gear

107m

Z

X

Y

Z

Motor

SECTION X-X

SECTION Y-Y
SolenoidIntermediate gear

Intermediate gear(parking brake mechanism
is equipped on the back)Lock pin
SECTION Z-Z(parking brake mechanism)

Fig. 2

Schematic of EMB unit

Table 1 Specification of EMB unit (for the rear wheels of 1500cc-class vehicles)
Item

Specification

Remarks

Service brake pressure

12 kN or greater

Hydraulic brake (cylinder dia. 33), hydraulic pressure 14 Mpa equiv.

Parking brake pressure

9.3 kN or greater

Pad µ=0.15, slope capability: 20% equiv.

Actuator size

107×62×84.5-30% compared to the conventional product (developed in 2011)(occupied volume)

Weight (excluding pad and mount)

2.8kg-0.4 kg compared to the conventional product (developed in 2011)

Table 2 Specification of motor
Item

Specification

Type

DC brushless

Size

46×52

Supply voltage

12V

Maximum output

140W(3) Power supply to the motor is stopped. Theintermediate gear rotates to the direction to reduce
the pressure due to the reaction force from thepad.
(4) The lock pin and the lock hole are engaged andthe rotation of the intermediate gear to thedirection of pressure reduction is restricted. Powerto the solenoid is switched off to end the operation.

2.2 Parking brake mechanism
Fig. 3 shows the schematic of the parking brakemechanism.
The parking brake mechanism consists of lock
holes on the intermediate gear, a solenoid mounted in
the motor housing, and a lock pin. The control flow of
the parking brake is described as follows using Fig. 4:[(b) Release](1) The motor is driven to the direction to increase thepressure.
(2) When friction applied to the lock pin is released,
the intermediate gear and the lock pin are
disengaged.
(3) The lock pin is pulled back by the return spring inthe solenoid.
(4) The motor is driven corresponding to the operationof the brake pedal (end of release action).

<Control-flow of parking brake>[(a) Operation](1) The pressure is added to the specified value by the
motor.
(2) The solenoid is activated to advance the lock pin.
-32-

*05E_03 14/07/31 13:48 ページ 4

Electro-mechanical Brake Unit with Parking Brake

Brake disc

Motor

Caliper

Solenoid
Lock pinLinear motion mechanism

Output gear

Intermediate gear

Input gear

Fig. 3 Schematic of parking brake

Intermediate
gear

Lock pin

Lock hole

PCD section(1) Operation command ON,
activate motor(2) Activate solenoid(3) Stop motor power supply(4) Lock pin engaged,
operation complete(a) Operation(1) Release command ON,
activate motor(2) Lock pin released(3) Lock pin receded(b) Release
Fig. 4 Control-flow of parking brake-33-(4) Release complete

*05E_03 14/07/31 13:48 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

3.2 Parking brake

3. Response

Fig. 6 shows the change of pressure when theparking brake is operated (released) when the service
brake pressure command is 0 kN (clearance: 0.1 mm).
To operate the parking brake the solenoid isactivated to advance the lock pin while the pressure
above a certain level is applied, as described in theprevious section. In that state, the pressure is
reduced and when the phases of a lock hole on theside of the intermediate gear and the lock pin match,the intermediate gear is locked (A in the figure). To
release the parking brake, the pressure is raised by
the motor first to release the friction applied to the lockpin (B in the figure) then the pressure is released.
The operation completes within 0.3 seconds from thecommand both for operating and releasing the parking
brake.
Similarly, it was confirmed that the operation
completes within approximately 0.3 seconds when theservice brake is operated (the brake pedal is pressed).

The evaluation result of the response, which is oneof the key characteristics of the unit, is described in
the following. An inverter dedicated to the electromechanical brake unit was designed and built for
conducting the evaluation and the pressure feedback
control where the load is detected by the load sensor
within the unit was configured.

3.1 Service brake

15

15

10

10

Pressure kN

Pressure kN

Fig. 5 shows an example of the evaluation results of
response of the service brake. It shows how the actualpressure tracks the pressure command of therectangular shape (0 kN⇔12 kN). The initial clearance
between the pad and disc is 0 mm. The result showsthat the response is approximately 30% improved overthe conventional products both for rise and drop ofpressure.

5
CommandEnhanced
product
Conventional
product000.1Time sec

5

CommandEnhanced
product
Conventional
product

00.20(a) Pressure rising (initial clearance: 0 mm)

0.1

0.2

Time sec(b) Pressure dropping

Fig. 5 Response of service brake

15

Operation

5

0
-0.1

Parking brake commandPressure
0

0.1

0.2

Time sec

0.3

0.4

Release0.5

Parking brake commandPressure
Operation

10

5

0
-0.1

0

(a) Parking operation (initial clearance: 0.1 mm)

0.1

0.2

Time sec

0.3(b) Parking release

Fig. 6 Response of parking brake

-34-

0.4

Release0.5

Parking brake command

10

Pressure kN

Pressure kN

15

*05E_03 14/07/31 13:48 ページ 6

Electro-mechanical Brake Unit with Parking Brake

4. Conclusion

5) Christain von Albrichsfeld， Jurgen Karner：Brake
System for Hybrid and Electric Vehicle，SAE Paper，
2009-01-1217(2009)．

In this article, we introduced a small electromechanical brake unit with a parking brake function
which is applicable for rear wheels of 1500 cc classvehicles. This unit achieved a reduced size andenhanced performance compared to the conventional
products by revisiting the internal design.
The enhanced performance of a brake is effective
not only for fuel economy of the vehicles but also forthe safety 10)-13); therefore, it is assumed that thedemand for the electro-mechanical brake will be
increased in the future. We will continue to develop an
electro-mechanical brake, improving the unit andcontrol technology for commercialization.

6) Yamasaki, Eguchi, Makino: Development of an
Electromechanical Brake, NTN TECHNICALREVIEW 75 (2007) 53-61.
7) Yamasaki, Eguchi, Makino: Actuator forElectromechanical Brake, NTN TECHNICALREVIEW 77 (2009) 40-44.
8) Tatsuya Yamasaki：Linear Motion Actuator forElectromechanical Brake，World Tribology
Congress 20099) Yamasaki: Development of Actuator forElectromechanical Brake, THE TRIBOLOGY, 285(2011) 34-36.
10) Tomohiko Adachi: The Electronic Stability Control
System, Journal of Society of Automotive
Engineers of Japan, Vol. 60, No.12 (2006) 28-33.

References
1) Yutaka Noguchi, Manabu Takegawa: Introductionof Electronic Parking Brake System, SUBARUTechnical Review, No. 36 (2009) 70-72.

11) Bo Cheng, Tadashi Hatano, Toshiya Hirose:
Effects of Brake Assistance Systems in Emergency
Situations, Proceedings of the Society ofAutomotive Engineers of Japan Annual Congress
No. 20065894 (2006).

2) Takuya Obata, Yukio Ohtani, Yoshinori Itou,
Shinnosuke Gotoh, Yuichi Koike: Development ofElectrically-Driven Intelligent Brake Actuator with
Regenerative Braking System, Proceedings of theSociety of Automotive Engineers of Japan Annual
Congress No. 20115172 (2011).

12) Yasuo Konshi, Masahito Hattori, Masakazu
Sugisawa, Michiharu Nishii: Development of
Hydraulic Brake Actuator for Active Brake Control,
Proceedings of the Society of Automotive
Engineers of Japan Annual Congress No.
No.9739075 (1997).

3) Naoto Okubo, Takashi Nishioka, Kohei Akamine,
Kunimichi Hatano: Development of Electric Servo
Brake System, HONDA R&D Technical Review,Vol. 25, No. 1 (2013) 47-51.

13) Hirohisa Tanaka, Hiroyuki Kodama, Yoshiki
Matsuzaki, Kaoru Tsubouchi: Innovative
Technologies of Braking System and Components,
Journal of Society of Automotive Engineers ofJapan, Vol 59, No. 1 (2005) 69-74.

4) Yasushi Aoki, Kenji Suzuki, Hiroshi Nakano,
Yoshikazu Nojiri: Development of Hydraulic Servo
Brake System for Hybrid Vehicle, HONDA R&D
Technical Review, Vol 18, No. 2 (2006) 60-66.

Photo of authors

Tatsuya YAMASAKI

Makoto MURAMATSU

Yui MASUDA

Chassis System Engineering,
EV Module Business HQ

Chassis System Engineering,
EV Module Business HQ

Control System Engineering,
EV Module Business HQ-35-

*06E_03 14/08/01 10:11 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Ball Screw Drive Module for Electric Hydraulic Brake

Hayato KAWAGUCHI*
Keisuke KAZUNO*

Recently, EV and Hybrid Vehicle percentage is getting higher at the
automotive market main caused by the low fuel consumption requirement
and CO2 emission reduction requirement. Those vehicles have theregenerative braking system, in order to obtain the vehicle running range
expansion and the highest electric efficiency. NTN have developed thecompact lightweight ball screw drive module for regeneration cooperative
brake system.
This report introduces the development result and the typical feature for it.

the maximum extent; however, the regenerative brakeand hydraulic brake must be used cooperatively to
instantaneously respond to the demanded braking
power that changes from time to time.
NTN has developed a high response, high thrust,compact and lightweight "ball screw drive module foran electric hydraulic brake” for a regenerationcooperative brake system which can instantaneously
and arbitrarily control a hydraulic brake.

1. Introduction

Braking force

The regenerative brake system used in hybrid andelectric vehicles uses a motor for driving the vehicles
as generators when the brake is applied in order torecover the kinetic energy as electrical energy.
However, the braking force by the rotational
resistance of the motor alone does not provide thenecessary braking power intended by the driver,therefore, the shortfall indicated by the shaded area in
Fig. 1 needs to be augmented by a hydraulic brake. In
order to recover as much electric energy as possible,
it is necessary to operate the regenerative brake to

2. Configuration of electric hydraulicbrake
The configuration of an electric hydraulic brakeused for a regeneration cooperative brake system isshown in Fig. 2. The mechanism to apply hydraulicpressure to the caliper of each wheel is as follows:
When a driver presses the brake pedal, the amountof pressure applied to the pedal is transmitted to ECU.
From this signal and the vehicle’s traveling informationsuch as the vehicle speed, ECU calculates thehydraulic pressure required for producing the
supplementary braking force by considering thebraking force required by the driver and producible
regenerative brake force. ECU drives the DC motor
based on this calculation result. The ball screw drive
module is operated by the DC motor pressing thepiston in the hydraulic cylinder to produce an

Braking power demanded by driver
Braking force by hydraulic brake

Braking force by motor(regenerative brake)

Start of braking

End

Fig. 1 Regenerative brake system**Chassis Engineering, Automotive Business HQ-36-

*06E_03 14/08/01 10:11 ページ 2

Ball Screw Drive Module for Electric Hydraulic Brake

appropriate pressure and the braking force is createdby each wheel caliper.
Adoption of the ball screw drive module realizes
precise control that responds to the demanded action
enabling recovery of additional energy even in thearea of low recovery efficiency such as the start andend of the braking action.

IN

Aluminum caseBolt

DC motor rotational torqueIdler gear

Ball screw

Output gear
SleeveDriver

DC motor

Liquid pressure
braking force

Ball screw
drive module

Brake pedal
operation of driver

ECU

OUT

Regenerative braking force

Thrust generation(pressing hydraulic cylinder)

InverterSupporting
bearings

Caliper

Hydraulic cylinder

Fig. 3 Cross section structure of ball screw drive module
Caliper

Motor
Idler gear + shell-shape needle bearings

Supporting bearings

Fig. 2 Constitution of electric hydraulic brake

3. Benefits of ball screw drive module
3.1 Compact, lightweight and simple design
The ball screw drive module is designed forcompactness and lightweight with a simple structurethat uses few components as shown in Fig. 3.
The rotational torque of the DC motor is transmitted
to the ball screw nut through the gear train for the
reducer for linear motion of the ball screw shaft.
The key components include a ball screw,
supporting bearings, sleeve as the rotation stopper,
gear train as reducer which transmits the rotational
torque from the motor, and aluminum case housing.
The ball screw drive module is a module productthat brought together not only the ball screw and the
supporting bearings, which are the key components
for converting the rotational torque of the DC motor toaxial thrust power, but also many of the NTN’s
proprietary technologies such as sintered materials.
The following is the description of the detailed
structure of the components:
Fig. 4 shows the details of the idler gear structure
for the reducer. The idler gear unit is inserted to the
supporting shaft with washers on both sides. Shellshape needle bearings are press fit inside the idle
gear to reduce the friction loss.
The joint of the aluminum case is sealed with aliquid gasket. The adoption of a liquid gasket thatSide washers

Fig. 4 Idle gear structure

forms an elastic and adhesive film enabled
followability to elastic deformation of the joint, as well
as a high level of airtightness.

3.2 Long-life, high-strength and environmentresistant specification(1) Long life and high-thrust power of ball screws
For long-life ball screws, the load capacity can be
improved by increasing the shaft diameter, ball size andnumber of balls. On the other hand, the smaller the leadthe more thrust power when the rotational torque isconverted to the axial thrust power by the ball screws.
However, if larger balls are used in excess of the size ofthe lead in the ball screw of normal design, the balls ofadjacent leads make contact to each other jeopardizing
the ball screw function.
-37-

*06E_03 14/08/01 10:11 ページ 3

NTN TECHNICAL REVIEW No.81（2013）(2) High-strength aluminum caseWe developed a long-life ball screw that allows ballsize in excess of the lead by taking note of the factthat the “balls are withdrawn to the deflector before
going on one circle” which is the characteristic of thedeflector type circulation and adopted it in this module.
Fig. 5 shows the schematic of products of
conventional design and long-life design.
Table 1 shows the specification comparison
between the ball screws of conventional design andlong-life design. By introducing larger balls andoptimizing the groove specification, we havesuccessfully developed compact, long-life, and highthrust power ball screws of double load capacity
compared to the conventional products without
increasing the nut length, outer diameter or leads.
In order to verify the durability, we conducted ballscrew shaft linear motion tests under specified loads
at a low temperature of -30˚C to a high temperature of
110˚C.
After these tests, we conducted performanceverification and disassembly inspection of the ballscrew drive module and the result indicated nosignificant changes in efficiency, airtightness, etc.
satisfying the set functions. No anomalies were foundin all the components including the ball screw andgears thus showing sufficient durability.

Aluminum die cast is adopted for the case of theball screw drive module. The benefits of die castmanufacturing include the ability to cast a thin andprecisely sized plate, superior mass productivity withshort cycle time, as well as high material strength due
to refined structure and formation of a chill layer (layer
with high internal density) on the surface by rapid
solidification.
The strength of the aluminum case of the ball screw
drive module was verified through FEM analysis inadvance and the maximum stress on the mostvulnerable part of the aluminum case was confirmed
before selecting the optimum thickness to secure asufficient safety factor.

(3) Environment resistanceThe ball screw drive modules are used in the brakesystem, which is one of the key vehicle components,
and therefore performance is required to be
maintained for a long time under diverse
environments. As it is exposed outside the body, theenvironment resistance was evaluated under hightemperature, high humidity, and salt spray tests.
After the tests, we conducted performanceverification and disassembly inspection of the ballscrew drive module and the results from all the tests
indicated no significant deterioration or change in sizethat would cause any issues in the function, thus
showing sufficient environment resistance.
Fig. 6 shows the case joint area after a cyclic
corrosion (salt spray) test. We verified that there wasno corrosion inside the aluminum case seal in the salt
spray test.
We also verified that there was no water immersion
inside the aluminum case seal after submerging theball screw drive module in high temperature water,
during the high-temperature submergence test.

Long life design

Conventional design

Increasedball size

Lead

Lead

Fig. 5 Schematic of long life ball screw

Table 1 Ball screw specifications comparison
Specification

Case joint

Long life design

Conventional design

Shaft diameter (mm)16

Lead (mm)4Ball diameter(/conventional design)

1

1.33

Dynamic load rating Ca(/conventional design)

1

2.06

Static load rating Coa(/conventional design)

1

2.06

No corrosion
inside the caseLiquid gasket

Fig. 6 Aluminum case after cyclic corrosion test-38-

*06E_03 14/08/01 10:11 ページ 4

Ball Screw Drive Module for Electric Hydraulic Brake

4. Conclusion
In this article, we introduced the ball screw drive
module adopted in the electric hydraulic brake.
The market for the compact and light-weight ballscrew drive module is expected to grow as vehicles
adopt more "by-wire" control and electric power.
NTN succeeded in the application of module
products consisting of mainly ball screws into key
vehicle components by this development.
We strive to continue development andenhancement of not only the ball screw technology but
also technology of peripheral components such ascases, reducers, and motors in order to achieve lightweight and improved functionality of module products.

References
1) Kazuno, et al.: Ball-screw Unit for Variable Valve
Event and Lift System, NTN TECHNICAL REVIEW
No. 75 (2007) 72-77.
2) Japan Aluminium Association, Aluminum Handbook(7th Edition)

Photo of authors

Hayato KAWAGUCHI

Keisuke KAZUNO

Chassis Engineering,
Automotive Business HQ

Chassis Engineering,
Automotive Business HQ-39-

*07E_03 14/08/01 10:31 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Low Torque Deep Groove Ball Bearings for EV and HEV

Tsuyoshi KODA*
Takahiro WAKUDA**
Tomohisa UOZUMI***In order to improve mileage per charge, various approaches for EV
and HEV parts to achieve low electricity consumption, weight
reduction, and downsizing are attempted. And also, for bearings to
be utilized in EV motors, further low torque performance is a strong
demand. This paper introduces High Speed and Low Torque Deep
Groove Ball Bearing for EV/HEV as Oil Lubrication Type and Low Torque Sealed Ball Bearing as
Grease Lubrication Type from NTN, and explains its features and performances.

existing bearings used for ordinary vehicles isapproximately 0.5×106 and any operation at higher
dmn conditions would cause problems such asheating up or deformation of cage, etc.
The developed product achieves low torqueoperation by optimization of internal specifications andimplementation of a newly developed cage. It can beused under high dmn conditions due to enhanced
cage strength.

1. Preface
In EV/HEV motors, two types of ball bearings,
namely, transmission oil lubricated and greaselubricated, are used. The oil lubricated type is used inhigh speed applications for small motors. On the other
hand, the grease lubricated type tends to be used inrelatively low speed applications for larger motors.
Two types of bearings have been developed tocontribute to low power consumption of motors,
namely, high speed/low torque, oil lubricated, deep
groove bearings and low torque, grease lubricated,sealed deep groove ball bearings; the following article
introduces them.

*1 Product of the diameter of bearing rolling element pitch circle

and rotational speed

2.2 Benefits of developed product
Table 1 shows a comparison of the conventional
product and developed product. The newly developed“combined type resin cage” is a cage that is made oftwo pieces of the same shape combined together asshown in Fig. 1.
Under conditions in excess of 0.5×106 dmn, astandard cage may deform by the centrifugal force
and make contact with the outer ring or rolling element
and/or be damaged by heat at the raceway. With thisdeveloped product, heating has been reduced by
optimizing internal specifications of the deep grooveball bearings, as well as implementation of a“combined type resin cage” which has little
deformation by centrifugal force. In addition,

2. High speed low torque deep grooveball bearings (oil lubrication)
2.1 Concept
In order to reduce electric power consumption of
EV/HEV, it is required to reduce the size and weight ofmotors. However, simple reduction of size would
cause a shortage of output torque of the motor, whichneeds to be supplemented by high speed rotation.
Under this high rotational speed condition, thebearings are required to operate with dmn value*1 ofapproximately 1.6×106. However, the dmn value of the***Powertrain Engineering, Automotive Business HQ***Drivetrain Engineering, Automotive Business HQ***Drive System Engineering, EV Module Business HQ-40-

*07E_03 14/08/01 10:31 ページ 2

Low Torque Deep Groove Ball Bearings for EV and HEV

Table 1 Comparison between ordinary and developed
Conventional productDeveloped productMaterial ofbearings

SUJ2

SUJ2

Material ofcage

SPCC, PA66

PA66

Conventional product (corrugated iron plate cage)
Conventional product (cantilever resin cage)
Developed product (combined type resin cage)

TorqueReduction of over 50%

Structure of
cage
Steel
plate

Combined type
resin

Cantilever
resin

500

1000

3000

5000

Rotational speed min-1

6000

Fig. 2 Torque measurement results
Flangegroove2) Durability
Table 3 shows the test conditions. When thedeveloped product was operated at 1.6×106 dmn for
1,000 hours, it did not show any seizure or damage ofthe cage. Fig. 3 shows the appearance of the bearing
components after the test. The developed productshowed no signs of damage, such as peeling andflaking, on the raceway of the inner/outer rings orrolling elements

FlangeFig. 1 Feature of combined type resin cage

temperature rise of the bearings in high speed rotation
has been lowered by reducing the oil inflow andstirring resistance with the flanges on the cage sides
and the flange grooves of inner/outer rings.

Table 3 Durability test conditions
Bearing size
Cage specification
Lubrication

6008
Developed combined type resin cageATF

2.3 Performance of developed product1) Rotational torque

Rotational speed

30,000 min-1

Load

Radial 1,000N

Table 2 shows the measurement conditions andFig. 2 shows the measurement results.
The developed product shows at least 50% lower
torque compared to the conventional products at a
rotational speed of 3,000 min-1 or more.

Operation time

Table 2 Torque measurement conditions
Bearing size

1,000hInner ring

Outer ring

Steel balls

Cage

Upgraded rated load based on 6909
Conventional corrugated iron plate cage

Cage specification

Conventional cantilever resin cage
Developed combined type resin cage

Rotational speed
Load

Max. 6,000 min-1
Radial 100N

Temperature

40˚C

Lubrication

ATF

Oil level

Fig. 3 Sample after durability testing

Lowest rolling element pitch circle diameter-41-

*07E_03 14/08/01 10:31 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

3) Strength of cage

3. Low torque sealed deep groove ballbearings (grease lubrication)

Fig. 4 shows the analysis result of the stressproduced on the cage by the centrifugal force under admn value of 1.6×106. The stress on the cage of thedeveloped product was verified to be under thematerial fatigue limit, exhibiting no problems. In
addition, the radial deformation was shown to be 1%
or less, a sufficiently small value.
One of the causes of the cage damage is stress due
to advance/delay of the rolling elements. The strengthof the cage has been verified by forcing
advance/delay of the rolling elements. This wasachieved by applying a moment with no lubrication.
Table 4 shows the test conditions and Fig. 5 shows
the test results under these conditions. The
conventional product broke at the pocket area in ashort time, however, the developed product continued
under the same condition, indicating superiority instrength compared to the conventional products.

3.1 Concept
There are 6 torque factors in grease-sealed deep
groove ball bearings as shown in Fig. 6. Factors (1),(2), (5) and (6) are determined by the internal
specifications of the bearings and can be obtained by
calculation.
The stirring torque due to grease viscosity, factor(3), is significantly affected by the grease properties.
In addition, the shear torque of the grease between
the cage and rolling elements, factor (4), is affected by
the configuration of the cage. For the developed
bearings, focus was given to factors (3) and (4) tolower the torque by improving the grease properties
and the cage configuration.
(1) Rolling viscosity torque on the raceway(2) Elastic deformation torque of the rolling element(3) Stirring torque due to the grease viscosity(4) Shear torque of grease between the cage androlling element(5) Torque due to the differential slip between therolling element and raceway rings(6) Torque due to the spin slip between the rolling
element and raceway rings

19
15

｛MPa｝

(1)

Fig. 4 Cage centrifugal force analysis results

(2)
(3)

Table 4 Cage strength test condition
6008

(4)

Conventional cantilever resin cage

(5)

Bearing size
Cage specification(6)

Developed combined type resin cage

Lubrication

No lubrication

Rotational speed

4,000 min-1

Moment

19.6N・mFig. 6 Torque factors

Time to breakage sec

800

Average: 600 sec(suspend)

600

3.2 Features
1) Adoption of newly developed grease
The newly developed grease for reducing the“stirring torque due to the grease viscosity” , factor (3),consists of a refined thickener and a base oil of low
viscosity. Table 5 shows the grease properties. As
shown in Fig. 7, the ability of the thickener to hold thebase oil was improved by refining the thickener
structure; this allowed reduction of the thickener
content and consequently, the resistance of thickener.
The resistance of the base oil was reduced by

400
200
0

Average: 36.3 sec

Conventional product(cantilever resin cage)

Developed product(combined type resin cage)

Fig. 5 Cage strength test results-42-

*07E_03 14/08/01 10:31 ページ 4

Low Torque Deep Groove Ball Bearings for EV and HEV

3.3 Performance of developed product1) Rotational torqueadoption of low-viscosity base oil. Both thesemeasures aided in reducing grease stirring of the
bearing rolling elements.

Fig. 10 shows the torque measurement result withthe measurement equipment and test conditions in
Fig. 9. It has been shown that the developed producthas at least 50% lower torque compared to theconventional products.

Table 5 Grease property table
Conventional
product

Developed
product

Thickener

Urea based

Urea based

Base oil

Synthetic oil

Synthetic oil

Base oil viscosity

40˚C

40.6

33.4

mm2/s

100˚C

7.1

5.8

Consistency

243

280

Starting
Low temperature
torque -40˚C mNm Rotating

350

250

80

70

Bearings to beevaluated

Load cell
Load

Static pressure table

Conventional
product

1µm

Developed
product

Bearing size

6203LLB

Rotational speed

4,000min-1

Load

1µm

Temperature
Greasing amountFig. 7 Thickener structure

200N
Room temperature
0.8g

Fig. 9 Torque measurement equipment and conditions

2) Adoption of newly developed cage
In order to reduce “shear torque of grease between
the cage and rolling element”, factor (4), it is importantto supply the minimum oil quantity without allowing
excessive grease inflow between the cage and rolling
element.
For the developed cage, a grease scraping
mechanism was added to the outer diameter of thecage pocket as shown in Fig. 8, to prevent excessive
grease inflow, and achieve lower torque.

TorqueReduction ofover 50%

Conventional
product

Developed
product

Fig. 10 Torque comparison results
Scraping grease off

Conventional
product

Developed
product

Fig. 8 Comparison between conventionaland developed products-43-

*07E_03 14/08/01 16:21 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

2) Grease life3) Strength of cage

Fig. 12 shows the result of evaluation for grease life
with the measurement equipment and test conditionsin Fig. 11. The developed product showedapproximately three times life when compared to theconventional products.

Fig. 13 shows the analysis result of the stressapplied to the cage by the centrifugal force. This stressis equivalent to the stress seen in the conventional
product, which does not exhibit any problems.
Cage strength has been verified by forcing
advance/delay of the rolling elements by applying
moment to the bearings. To maximize the stress on thecage, the cage pocket clearance was set to the lower
limit. Fig. 14 shows the moment load tester and test
conditions and Fig. 15 shows the strength test results.
It has been verified that there is no problem with thecage durability of the developed product.

Bearings to beevaluated

Load

Bearings to beevaluated｛MPa｝
Cartridge heater

Bearing size

6204LLB

Rotational speed

10,000min-1

Load2,000N

Temperature

150˚C

Greasing amount

1.0g

20.8

19.5

Conventional productFig. 11 Grease life evaluation tester and conditions

Developed productFig. 13 Cage centrifugal force analysis results
200
Test sample

Grease life h

150

100

Load direction

50

0Conventional
product

Developed
product

Cage specification Developed cage with the lowest pocket clearance
Grease typeFig. 12 Grease life evaluation test results

Developed grease

Rotational speed

5,000min-1

Moment11.8N・mTest time

89 h (loading times: 1×107)

Fig. 14 Moment load tester and conditions

No anomalies such as breakage and cracks are observed
Fig. 15 Moment load test results-44-

*07E_03 14/08/01 10:31 ページ 6

Low Torque Deep Groove Ball Bearings for EV and HEV

4) Acoustic property

4. Summary

As the drive power changes from engine to motor,
the requirement for noise reduction of the bearings will
also become more vital. The developed bearings have
improved acoustic properties compared to theconventional bearings. This has been achieved byrefining the thickener structure of the grease. Fig. 16
shows the acoustic tester and test conditions and Fig.
17 shows the acoustic test results.
It has been shown that the developed product issuperior in reducing noise as compared to theconventional bearings.

The developed product introduced in this article has
the characteristics as shown below. The developed
products will be promoted to contribute to compact
and power-efficient EV/HEV motors which are
expected to expand in the market.

[Low torque deep groove ball bearings for EV and HEV]High speed rotation
Can be used under conditions with a dmn value of
1.6×106[In case the bearing pitch diameter is 54.0 mm,
rotational speed of 30,000 min-1]
Running torqueOver 50% reduction compared to conventional
products

Microphone
Diaphragm

45˚[Low torque sealed deep groove ball bearings]

Bearing size

6203LLB

Rotational speed

1,800min-1

Axial load

20N

Temperature

Room temperature

Running torqueOver 50% reduction compared to conventional
products
Acoustic
Equivalent or better than our conventional
products
Life
Equivalent or better than our conventional
products

Complied with JIS B 1548

Fig. 16 Acoustic test conditions

Acoustic pressure value

Reduction of
approx. 3 dBA

References

Conventional
product

1) Proceedings of the Japan Society for Precision
Engineering Autumn Meeting in 2001, Agitation
torque that occurs with bath lubrication of rolling
bearings (report 1 and 2)

Developed
product2) Japanese Society of Tribologists, Tribology
Handbook, Yokendo, 2001, 173

Fig. 17 Acoustic test results

Photo of authors

Tsuyoshi KODA

Takahiro WAKUDA

Tomohisa UOZUMIPowertrain Engineering
Automotive Business HQ

Drivetrain Engineering
Automotive Business HQ

Drive System Engineering
EV Module Business HQ-45-

*08E_03 14/08/01 10:33 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Article ]The Reactor Core for HEV Boost Converter

Takayuki ODA*
Shinji MIYAZAKI*
Eiichirou SHIMAZU**
Nippon Kagaku Yakin Co.,Ltd., as one of the NTN group companies,
develops various types of magnetic material products1). The major
product is the Amorphous core which has characteristics of highsaturation magnetic flux-density, low core loss and superior frequency
characteristics.
This article introduces the Amorphous reactor core to meet high current and compact for Hybrid
electric vehicle.

In this article, we are introducing the amorphousalloy-based cores with high saturation magnetic fluxdensity, low iron loss, and superior frequency
characteristics.

1. Introduction
Vehicles are equipped with batteries and manyelectric parts which require their respective suitable
voltages. Therefore, DC-DC converters are used toconvert DC voltages 2). Reactors in the DC-DC
converters are electromagnetic components used toconvert electric energy to magnetic energy andstore/release it in the switching circuits. They also
provide voltage boosting/dropping and current
smoothing. Fig. 1 shows the key power control
components in hybrid vehicles (HEVs) and a typicalreactor configuration used in the booster circuit.
As the DC-DC converters for HEVs supply power tothe drive motor, which requires a large output, reactor
cores are required to support large current of several
tens to hundreds of amperes. In addition, demand forcompact reactors is strong since they are mounted
onto vehicles. Since higher switching frequency iseffective for achieving more compact reactors, it isnecessary to address this technology.
Nippon Kagaku Yakin Co., Ltd., as one of the NTN
group companies, develops core materials mainly
using amorphous alloys to address these challenges.
These materials provide more compact reactors
compared to widely used ferrite materials. It is also
possible to reduce iron loss compared to the Fe-Si
based alloy materials.

EngineBattery
Generator

Motor

Inverter

Booster
circuit

Reactora) Reactor application in electric vehicleb) Reactor configuration

Fig. 1 The reactor for booster circuit**Engineering, Nippon Kagaku Yakin Co., Ltd.
**Advanced Technology R&D Center-46-

*08E_03 14/08/01 10:33 ページ 2

The Reactor Core for HEV Boost Converter

superior soft magnetic material because of its highpermeability and low iron loss. Unlike ordinary soft
magnetic materials, amorphous alloys have a coercive
force since there is no crystal grain boundary which
prevents movement of the magnetic domain wall andhas no magnetic anisotropy. It is also extremely hard
and resistant to corrosion.
Table 1 shows comparison of properties with othermagnetic materials.

2. Types and characteristics ofamorphous alloy cores
2.1 Types and characteristics of the developed
core materials
While an amorphous alloy has excellent magnetic
properties, mass production is difficult due to its
extreme hardness. Mass production also causes
wear on the mold during compression or injection
molding and a high burden on the facilities. We have
solved this difficult problem by developing our own
materials and proprietary manufacturing technologies.
These include improving the binder composition
blended in the amorphous alloy powders, improving
manufacturing facilities and mold structures,
optimizing molding conditions, and many more.
Fig. 2 shows typical examples of cores developedby Nippon Kagaku Yakin Co., Ltd.
Two types of materials, “AL60” for compression
molding and “AS10” for injection molding, were newly
developed. Both materials exhibit a good frequency
response by maintaining the magnetic properties up to
the higher frequencies. AL60 has a high relative
permeability and small iron loss. On the other hand,
AS10 has a lower relative permeability and theinductance drop on high current is smaller. AS10 also
has a higher degree of freedom for design since ituses injection molding. Inductance is the capability ofa coil to generate flux from current. It is important forcoils, such as reactors, to maintain their magnitude
regardless of the current values.
By taking advantage of these properties, Nippon
Kagaku Yakin Co., Ltd. is designing and proposing
cores using either type of material or combining bothmaterials to create composite cores. These designs
are selected based on the specifications such as thecurrent values and sizes required by customers.

a) Compression molding core AL60b) Injection molding core AS10

2.2 Properties of amorphous alloys
Fig. 3 shows a SEM photograph of amorphous alloy
powders. An amorphous alloy is a non-crystalline
metal with a disordered atomic structure. This is
produced by rapid cooling of over 1,000,000 °C/sec
from a melted state. Since the alloy does not havecrystalline structure, it shows the following properties 3):

AS10

AL60
AL60

・Isotropic physical property
・No heterogeneous property such as a crystal
grain boundary, twin crystal, lamination defect,
or dislocation in the structure
・Unlikely to deform because of the no slip planeof crystalc) Composite core (AL60+AS10)

Based on these properties, an amorphous alloy is a

Fig. 2 Core of various processes-47-

*08E_03 14/08/01 10:33 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

On the other hand, AS10 for injection molding uses
polyphenylenesulfide (PPS) which has excellent heat
resistance as a base material. When it is used withlarge current, such as in the reactor application for
HEVs, the coil and core will be significantly heated. By
using PPS as the base material, they can withstandup to 200°C. Typically when the mold is highly filledwith magnetic particles, the fluidity of the material
during molding is reduced resulting in unstable fillingof the mold. However, with AS10, both high filling andhigh fluidity are achieved by optimizing the grain sizeand shape of the amorphous alloy powders andadditive during kneading.

Fig. 3 Amorphous magnetic powder

3.2 Manufacturing technology

Table 1 Comparison of material characteristics 4), 5)
Permeability
Magnetic flux density
Iron loss
Coercive force
Workability
Wear resistanceCorrosion resistance

Amorphous alloy

Fe-Si alloy

Ferrite

○
○
○
○
×
○
○

○
○
△
△
△
×
×

○
×
○
○
△
△
○

○：Superior△：Slightly inferior

Vickers hardness of an amorphous alloy is over
1,000, which is approximately 10 times harder than aferrite material.
With compression molding, the method used for
manufacturing AL60 cores, the material powders arefilled in the mold, compressed, and calcinated.
Compression molding of ordinary metal particles
requires pressure of approximately 500 MPa.
Amorphous alloy particles require a significantly
higher pressure of approximately 2,000 MPa since thealloy is extremely hard and lacks a plastic deformation
property. Therefore, slight misalignment or inclinationin molding components and/or the press machine may
generate extremely high partial pressure. This causes
early wear of the mold or a large load on the auxiliary
facilities. We have established ultra-high pressure
molding technology improving precision of the pressmachine by providing a linear guide and achieving ahigh precision of 1/1000 mm for the mold.
The injection molding method for cores made ofAS10 is required to provide measures against wear of
the kneading machine, molding machine, and mold.
This is necessary since hard amorphous alloy
powders act as abrasives during kneading, pelletizing,and injection molding. We have established mass
production technology by optimizing the component
materials, surface treatment, mold structure, andmolding conditions.

×：Inferior

3. Developed technologies
3.1 Material technology
Amorphous alloys starts crystallizing when heatedto around 500°C and starts losing the optimal
properties of amorphous materials. To prevent this,the binder used as the compression molding material
needs to have hardening property at low calcination
temperature after being molded.
It also needs to withstand the annealing process ataround 450°C which is applied for magnetic property
enhancement after molding and calcination.
Thermosetting resin, used as a binder for bondmagnets, hardens at low temperatures. However, arelatively high amount of the resin content is required
to secure the product strength. This reduces the filling
ratio of magnetic particles making it difficult tomaintain a high magnetic property required for reactor
cores of HEVs. Also, the low heat tolerance of thebinder content prevents the annealing process.
Therefore, we developed a proprietary bindermaterial for AL60 that enables molding and calcination
at low temperatures with a small quantity of binder
mix. It also allows the annealing process to take placeat 450°C.
-48-

*08E_03 14/08/01 10:33 ページ 4

The Reactor Core for HEV Boost Converter

Magnetic
saturation

4. Performance of developed core
4.1 Large current characteristics
400

Magnetic flux density mT

Ferrite cores consisting of iron oxide are generally
used for reactors in low power output equipment that
produce less than several hundred watts of output.
The ferrite cores have extremely high relative
permeability of several thousands, compared toamorphous cores with several tens. They also have
low saturation magnetic flux density.
In a large current application such as a DC-DC
converter for HEVs, the magnetic energy becomeslarge and the flux through the cores tends tomagnetically saturate. At magnetic saturation,inductance of the coil suddenly drops and it is difficult
to maintain the properties of the reactor. Therefore, inorder to prevent magnetic saturation of ferrite coreswith low saturation magnetic flux density, it isnecessary to increase the cross sectional area toreduce generated flux density. This is not suitable forvehicle components which are required to be morecompact.
Fig. 4 shows the relation between flux density andmagnetic field intensity of the AL60 and AS10
amorphous alloy cores, an ordinary ferrite core, and a
Fe-Si based compression molding core.
It shows that the ferrite core and the Fe-Si based
compression molding core have a steep slope. The
ferrite core also saturates at a small field intensity andthe flux density does not increase. Since the AL60
compression molding core and AS10 injection moldingcore do not saturate, they can be applied to DC-DC
converters for HEVs, which require a large current.
The change rate of inductance over the change of
DC current, called the DC superimposition
characteristic, is small with the amorphous cores. As
shown in Fig. 5, inductance of the ferrite cores drops
around 80% at 50A. However, inductance of AS10
only drops around 10%, even at 100A, showing an
excellent DC superimposition characteristic.

350
300
250
200

Ferrite core150

AS10 injection molding core100

AL60 compression molding core50

Fe-Si based compression
molding core

0

0

1000

2000

3000

4000

5000

6000

Magnetic field intensity A/m

7000

Fig. 4 B-H characteristics

Change rate of inductance %120
AS10 injection molding core
AL60 compression molding core
Fe-Si based compression molding core
Ferrite core

100
80
60
40
20
0
0

50

100

150

Current A

200

250

300

Fig. 5 DC superimposition characteristics

Change rate of inductance %120

4.2 High frequency properties
Reactors for HEVs are required to support large
current and high switching frequencies within a small
size requirement.
Fig. 6 shows the relation between the frequency
and the inductance change rate for different cores.
The inductance drops at several kHz with the ferrite
core and several hundred kHz with the Fe-Si based
compression molding core. On the other hand, theinductance of the developed amorphous cores stayed
constant, even in the high frequency range up toaround 1000 kHz.

100
80
AS10 injection
molding core60

AL60 compression
molding core40

Fe-Si based compression
molding core

20
0Ferrite core

1

10

100

Frequency kHz

Fig. 6 Frequency characteristics

-49-

1000

*08E_03 14/08/01 10:48 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

In the large current applications, Fe-Si based
compression molding cores are frequently adopted,
which have a large saturation magnetic flux density.
However, Fe-Si based compression molding cores
have relatively larger iron loss compared to ferrite
cores (Fig. 7). The amorphous alloy of Nippon Kagaku
Yakin Co., Ltd. has reduced iron loss equivalent to theferrite cores by applying a special insulation coating
treatment.
As stated above, the developed AL60 and AS10
amorphous alloy cores are capable of responding to
the demand of large current and high frequencies withsmall volume. They have a high saturation magnetic
flux density, excellent frequency properties, low ironloss, and DC superimposition characteristics.

6000

Magnetic flux
density distribution T

AS10 injection molding core

8.00e - 01
7.50e - 01

AL60 compression molding core

5000

Iron loss kW/m3

may cause abnormal heating. This information isprovided by analysis of the current on the coil,intensity of generated magnetic field, and magnetic
flux density.
NTN is also conducting thermo and electromagnetic
analysis through the finite element method. Fig. 9shows an example of the temperature distribution
analysis of the reactor. By optimizing conditions foranalysis, such as the heat transfer coefficient,temperature distribution of the product in operation
can be precisely analyzed. This analysis iscontributing to the verification and countermeasuresfor heat resistance and radiation.

7.00e - 01
6.50e - 01

Fe-Si based compression
molding core

4000

6.00e - 01
5.50e - 01

Ferrite core

5.00e - 01
4.50e - 01

3000

4.00e - 01
3.50e - 01
3.00e - 01

2000

2.50e - 01
2.00e - 01
1.50e - 01

1000
01.00e - 01
0.50e - 02
0.00e - 00

0

20

40

60

80

100

120

Fig. 8 Magnetic flux density vector distribution of a reactor

Frequency kHz

Fig. 7 Iron loss characteristics (50mT)
AS10

AL60
Temperature

5. Optimization design by simulation

Contour plot：deg C
35.0

For restricted size and high power output of thereactor core, it is necessary to increase the drive
frequency. The AC resistance of the coil and iron lossof the core increase as the frequency increases.
Therefore, the design needs to consider radiation ofheat from copper loss of the coil and iron loss of thecore.
Since it takes time to conduct experiments and trial
manufacturing for this type of design, electromagnetic
field analysis simulation is used. It has been provento be an effective design tool for faster development
and estimation of properties difficult to verify inexperimentation.
Fig. 8 shows an example of magnetic flux density
vector distribution based on electromagnetic analysis.
We are utilizing this to optimize the design of the core
structure by confirming magnetic circuit configuration
and uneven magnetic flux density distribution which

31.8

26.3
23.1

Coil

20.0
Max.：34.14
Min.：20.00

20°C cooling pad

Heated temperature Coil
(°C)
Core

Experimental
data

Analyzed
data

40

34

42

35

Fig. 9 The temperature distribution analysis of reactor-50-

*08E_03 14/08/01 10:33 ページ 6

The Reactor Core for HEV Boost Converter

6. Summary
Nippon Kagaku Yakin Co., Ltd. developed anamorphous alloy with excellent magnetic properties:<Benefits of the developed core>(1) High saturation magnetic flux density and low
degradation of properties, even with large
current(2) Low degradation of properties at highfrequencies(3) Low iron loss
We are planning to propose our reactor cores for
HEVs to our customers, taking full advantage of the
optimized design by electromagnetic and thermal
analysis and proprietary manufacturing technologies.
This technology will contribute to the improvement ofdriving performance of HEVs and EVs.

References
1) Takuji, Harano, Shinji Miyazaki, Hajime Katsuura:
Introduction of Magnetic Material Products, NTN
TECHNICAL REVIEW No. 80 (2012) 87-91
2) Masayuki Morimoto, Kiyoshi Kanamura et al.:Latest Control Technology for Electric Vehicles,
NTS (2011) 78
3) Japan Electronics and Information Technology
Industries Association: Soft Magnetic Metal
Materials
4) The Society of Non-Traditional Technology:
Magnetic materials of the new era5) Hirota: The Magnetics Society of Japan, Bulletin of
the 4th Topical Symposium

Photo of authors

Takayuki ODA

Shinji MIYAZAKI

Eiichirou SHIMAZU

Engineering,
Nippon Kagaku Yakin Co., Ltd.

Engineering,
Nippon Kagaku Yakin Co., Ltd.

Advanced Technology
R&D Center-51-

*09E_*09 14/08/01 10:38 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Article ]Hub Bearing
with an Integrated High-Resolution Rotation Sensor

Kentaro NISHIKAWA*
Toru TAKAHASHI**
Christophe DURET***
Wheel Speed Sensor is one of the most important sensors on a vehicle andused not only for ABS but also for various control systems. We have developed
a new sensor device for high-resolution rotation sensing application and now thesensor can be integrated on hub bearings. In this article, we present the functionof the high-resolution rotation sensor hub bearings along with some of the novel
data measured on the test vehicle.

1. Introduction
Wheel speed sensors are critical components forvehicle safety control and are widely used not only for
anti-lock brake systems (ABS), but also for various
control systems such as electronic stability control(ESC).
Sensors have evolved from the passive sensor
method consisting of a pick-up coil and a pulsar ring to
the active sensor method that employs magnetic
sensors such as hall elements and MR elements
enabling consistent detection of rotational speed frompark to high speed driving.
NTN is promoting unit development where sensors
are also integrated into the seal of the hub bearings
with magnetic encoders and has commercialized ABS
sensor integrated hub bearings (shown in Fig. 1), andsealed sensor integrated hub bearings with enhanced
reliability 1).
While bearings with rotation sensors (shown in Fig.
2) are used in industrial machines and are equippedwith output resolution of several tens of pulses per
rotation, the demand for higher resolution is increasingalong with the evolution of control systems. However,increasing resolution by refining the magnetization
pitch of the magnetic encoder causes reduction of themagnetic force. Therefore, NTN-SNR has been
pursuing the improvement of the sensor capabilities
and has developed a high-resolution sensor withupconversion detection capability (MPS40S) 2). With

Fig. 1 ABS sensor integrated HUB bearings

Magnetic sensor(hall IC)

Basic configurationof sensor

Magnetic
field
Bz
Hall IC
Magnetic encoder(NS poles alternately magnetized)

Fig. 2 Rotation Sensor Bearing***Chassis Engineering, Automotive Business HQ***New Product Development R&D Center***NTN-SNR ROULEMENTS Research & Innovation Mechatronics-52-

*09E_*09 14/08/01 16:23 ページ 2

Hub Bearing with an Integrated High-Resolution Rotation Sensor

this technology the detection of rotation is now
possible with a resolution that is 40 times higher thanthe conventional method, even if the sensor is
combined with conventional magnetic encoders. NTNSNR has also developed a tunnel magnetoresistanceeffect (TMR) element with high sensitivity as the next
generation magnet sensor and is expected to be usedin the vehicle application as a magnet sensor operablewith a large air gap 3), 4).
The developed high-resolution sensor (MPS40S)
complies with the environmental resistancespecification required for vehicles and NTN is pursuing
its application in the hub bearings. In this article, we
provide an overview of hub bearings integrated withhigh-resolution rotation sensors and introduce someexamples of analyses of high resolution signals
measured with the sensor-equipped vehicle.

For drive wheel

For idler wheel

Fig. 3 High-resolution rotation sensor HUB bearings

High-resolution sensor(IC/Processing circuits)
Resin mold

2. Overview of high-resolution rotation
sensorsSeal

2.1 Structure/configuration

Magnetic encoder

Seal withmagnetic encoder

Fig. 3 shows pictures of high-resolution rotation
sensors integrated hub bearings. For driving wheels,
the sensor is molded with resin together with the fixing
core grid and press fit onto the stationary ring (outerring) of the hub bearings (shown in Fig. 4). For idler
wheels, the sensor is fixed inside the water-resistantcap.
In either case, sensor assembly and adjustment
process on the vehicle can be omitted since theposition between the magnetic encoder and thesensor is controlled by integrating the sensors onto
the hub bearings.

Fig. 4 Integration of the high-resolution sensor

Table 1 Electrical specification of the rotation sensor
Item
Supply voltage

Value
Vdd[V]

Consumption current Ic [mA]
VH[V]Output voltage
VL[V]

5±0.5
40 or less
Vdd×0.8
0.4 or lessMax. output current Iout [mA]
Start time

2.2 Electric specification/output signal

Number of output pulses

Mainstream ABS sensors for vehicles use 2-wireinterface supplying 9-15V battery power, detecting 14
mA or 7 mA operating current that changes with therotation. On the other hand, the developed highresolution sensor uses 5V of supply voltage andoutputs a digital voltage signal on the independent
signal line. Therefore, it can be connected to thestandard digital input terminals that vehicle electric
control units (ECUs) have without any conversion
circuits.
Table 1 shows key electrical specifications of highresolution sensor integrated hub bearings. Since thesensor is equipped with output of A/B phases similar
to the ordinary rotary encoder output signals, it iscapable of detection of the rotational direction, quadedge evaluation processing using A/B phase edges,
etc. These electrical specifications comply with thesensor IC specifications 5).

Multiplication
number settingmx

I/O connection

Vdd
GND
VoA
VoB
VABS

±15
500µs or lessNumber of encoder pole pairs (PP)×multiplication number (mx)x40, x20, x10, x5,x32, x16, x8, x4, x2
Power input
GND
A-phase output
B-phase output
ABS signal output ** Optional (output signal without multiplication)

2.3 Resolution and accuracy of rotational speed
detection
The high-resolution rotation sensor realizes themultiplication process shown in Fig. 5. This means
the duration of one pair of magnetic poles (cycle) is
electrically interpolated to generate pulses with cyclesof 1/(multiplication number mx). The interpolated
output signals are superimposed with slight errors
depending on the distortion of the input magnetic field-53-

*09E_*09 14/08/01 10:38 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

frequency even at ultra-low speed(2) A sufficient amount of data can be obtained evenfrom a short traveling distance(3) Rotation immediately after start or immediately
before stop can be accurately detected
Table 3 shows examples of high-resolution rotation
sensor signal applications. In the following sections,
we introduce run data of the test vehicle.

and characteristics of multiplication circuits. Therefore,
the pulse width slightly fluctuates, synchronized withthe cycles of the magnetic poles of the magnetic
encoder even when the rotation is at a constant
speed. The magnitude of fluctuation is 0.5% or lessover the cycle of one pole pair and has acharacteristic of repeating along with the cycle of themagnetic poles. Therefore, when the rotational speed
is detected using the multiplied pulses, the impact ofmultiplication error can be effectively reduced byapplying a filtering process that uses pulse data fromthe number of pulses equivalent to one magnetic polepair cycle.

OutputHall element

High-resolution sensor

3.2 Test vehicle with the sensor
We drove the test vehicle with the high-resolution
rotation sensor. Table 4 shows the specification of thesensor and the test vehicle.
The resolution of the rotation pulses were set to 480
pulses/rotation, which is 10 times higher than theconventional product. The sensors were mounted on
the four wheels of the rear wheel drive vehicle tocheck the difference of rotation behavior between thedrive wheels and the idler wheels.

Highly precise detection ofmagnetic field distribution bymultiple hall elements
Multiplication processing:
Precise interpolation between
magnetic poles

Magnetic encoder

48 Pole pairs

Conventional output

Resolution 7.5˚High-resolution output

Resolution 0.18˚

Table 2 Comparison of output resolutions
Number of
output pulses(per rotation)
Conventional ABS sensor
High-resolution
rotation sensor

Fig. 5 Interpolated signal of the developed sensor( 40 times at most)

Resolution of Interval of output
pulses whenmovement
driving at 1 km/h
detection

48

40mm

150ms

1,920

1mm

3.7ms

Tire radius of 305 mm assumed

Table 3 Application of high-resolution signals
Vehicle condition to be evaluated

3. Output signal of high-resolution
wheel speed sensor

Tire rotation at
1 km/h or less

3.1 Application of high-resolution rotational signal

Determination of rotational direction

Hill hold

Drive/brake torqueImproved TCS*

Higher accuracy and response fordifference of rotation speed ofright/left wheels

Improved ESC**

Drive/brake

Conventional wheel speed sensors output around48 pulses per rotation. Converting this to the running
distance, it is equivalent to the resolution of one pulse
per around 40 mm.
The high-resolution rotation sensor provides
resolution of one pulse per 1 mm of running distance.
Table 2 shows the output pulse cycle at anextremely low traveling speed of 1 km/h. The interval
between output pulses of the conventional sensor is
150 ms meaning it takes a maximum of 150 ms before
detecting rotation. However, using output signals of
the high-resolution rotation sensor, the detection iscompleted in 3.7 ms. Therefore, even at a ultra-low
speed it is possible to ensure a sufficiently fast data
update rate.
Thus, the benefit of using a high-resolution rotation
sensor is its enhanced space and time resolution.
The following describes its benefits compared toconventional products.
(1) Rotational signals can be obtained with sufficient

Application
Automatic driving/parking
Hill hold

Tire condition

Change of tire pressure

Indirect TPMS***

Change of shape/
abnormal wear

Vehicle control in general
Safety alarm device

Dry/wet( µ estimation)

Vehicle control in general

Road condition*Traction Control System**Electronic Stability Control***Tire Pressure Monitoring System

Table 4 Specification of the test vehicle
Item
Vehicle Drive
Tire

SensorSize

Specification

Remarks

RWD

Rear Wheel Drive

225/55R17

Installation location

All wheels

Common to four wheels

Magnetic encoder

48 pole pair/rotation

90-degree phasedifference output

10 times

480 pulses/rotation

Multiplication number
Output signal

A-phase, B-phase

Measure- Pulse detection method Measurement of time between pulse edges
mentSampling frequency
80MHz

-54-

*09E_*09 14/08/01 10:38 ページ 4

Hub Bearing with an Integrated High-Resolution Rotation Sensor

3.3 Running data (1)-Estimation of the road conditions-400

A method of estimating road conditions is being
proposed by analyzing frequency response of speed
variation components extracted from the wheel
rotation signals of running vehicles 6), 7). Fig. 6 shows
an example of estimating tire-road surface conditions
from the variation of rotational speed. In this kind ofanalytical method from frequency response, dataneeds to be extracted under the condition of constant
speed since the spectrum resolution degrades as thetraveling speed changes. It is not easy to extract
sufficient varying speed information using theconventional sensors when vehicles run at low speed.

300

Frequency Hz

350
Road surface with sand

250
200
150
100
50
0

0

50

100

150

200

Run m

250

300

Fig. 7 Time-domain spectral diagram(Vibration frequency (Hz) in vertical axis)
50

T(1)

T(2)

T(3)

T(959)

T(960)

Wavenumber 1/m

OutputsignalOne tire rotation (480 pulses)
Measurementof pulses

T (n)Calculationof speedω (n)Analysis of
speedvariation

Estimation oftire-roadconditions

40
30
20
10
0

Fig. 6 Signal processing using the rotation signal

Road surface with sand

0

50

100

150

200

Run m

250

300

Fig. 8 Spatial-domain spectral diagram(Spatial frequency -wavenumber (1/m) in vertical axis)

As a way to ease this extraction conditions, at NTNSNR we are also exploring a method to estimate roadconditions by order frequency analysis of the speed
variation components adopting a method to sample
speed on the basis of angle-of-rotation using the highresolution sensor signals 8). In the following, we
describe an example of run data processing with thetest vehicle, to show how to extract variation ofrotational speed.
After starting and accelerating for about 30m, rollingon the asphalt road at 25 to 30 km/h, we drove onto
the asphalt road with fine sand on the surface threetimes. Fig. 7 and 8 show the frequency analysis
results of the speed variation components extracted
from the rotational pulse signals of the front rightwheel.
These figures show the results of FFT processusing speed information at the respective times for
one rotation of the wheel, sequentially displaying thespectrum calculated at the interval equivalent to 1/10of the wheel rotation which is about 0.2 m of travelingwith the traveling distance on the horizontal axis.
Fig. 7 uses frequency (vibration) for the vertical axis
and Fig. 8 uses spatial frequency (wave number) for
the vertical axis. The color bar on the right of thegraphs shows the correlation between the strengthand color and indicates that the colors on the upperarea of the color bar are stronger. Fig. 7 only plots thecomponents up to the 100th order, therefore,
frequencies over those components are left blank.
The speed variation spectrum of the running vehicle
shows changes in the areas corresponding to the roadsurface with fine sand. In these areas, it is revealed
that the speed variation is increased and the variationis spread out to the high frequency range. Because
high-resolution rotational pulses are used, frequency
components higher than the conventional methods
were extracted. It also shows that higher ordercomponents can be extracted even when the drive
speed is not constant since sufficient data can beobtained even on a short drive.

3.4 Running data (2) -Detection of rotation at
acceleration/decelerationDuring initial acceleration and acceleration/
deceleration on slippery roads, vehicles are likely to
become unstable because of the wheel slippage. Useof high-resolution rotation sensors enables detailed
observations of the wheel behavior at extremely low
speed for early detection of slippage or determinationof whether the wheel is locked immediately before-55-

*09E_*09 14/08/01 10:38 ページ 5

NTN TECHNICAL REVIEW No.81（2013）resolution rotation sensor, it is possible to provide
effective information for traction control on low μ
roads (slippery roads) which requires refined torquecontrol, etc.

stopping.
As an example, Fig. 9 shows the rotation speed
data of the front wheel (idler wheel) when a car stops
at a traffic light with the brake applied and starts rolling
after releasing the brake.
A variation of rotation is observed at the area
indicated by an arrow. This shows that the vehicle
stopped with load applied by braking and strain wasgiven to the suspension. The observation then reveals
that when the brake is released, the front wheel was
pushed forward and started rotating. On the other
hand, from the output signals of the ABS sensor
collected at the same time, this kind of observation
was not obtained and the detection of rotation wasapproximately 0.28 seconds late.
Fig. 10 shows a behavioral data of the rear wheel(drive wheel) when a car begins accelerating on aslippery road as another example. At around 0.65
seconds, a slippage is observed and the highresolution rotation sensor signals display a smooth
curve of how rotation speed rapidly increases. The slip
is detected about 0.05 seconds earlier than the speed
data from the ABS signals and a sufficient number ofdata is also acquired for calculating acceleration fromthe rate of speed variation. Thereforewith the high-

The rotation sensor signals reflect the roadconditions.
While the example described in section 3.3 used
analysis results of variation of rotation speed observed
for a certain duration, the example in the following
uses a single input to detect road disturbances.
Fig. 11 shows a variation of rotation speed of four
wheels when a vehicle passed a joint of the roadsurface while driving on a bridge at the speed of 20
km/h. An oscillatory waveform of 15 to 35 Hz is
observed at the joint, however, the waveforms andvibration frequency are different between the idler anddrive wheels. It is considered that the road conditions
and the driving force to the wheels are reflected to therotation speed signals.

6.4
6.35
6.3
6.25
6.2
6.15
6.1
6.05
6
5.95

28.8

High-Resolution Sensor
ABS Sensor

0.20

High-resolution sensor

Speed m/s

Rotation speed s-1

0.25

3.5 Running data (3)-Observation of non-repetitive phenomena-

0.15
0.10

0
0

0.1

0.2

0.3

0.4

0.5

Time s

0.6

0.7

28.8

6.3
6.25
6.2
6.15
6.1
6.05
6
5.95
5.9

28.8

ABS sensor

0.05

6.35
6.3
6.25
6.2
6.15
6.1
6.05
6
5.95

0.8

0.9

6.3
6.25
6.2
6.15
6.1
6.05
6
5.95
5.9

1.0

28.8

x10
ABS

29.0

29.2

29.0

29.2

29.0

29.2

29.0

29.2

Fig. 9 Rotation signal after releasing brake

Time s

29.6

29.8

29.4

29.6

29.8

29.4

29.6

29.8

29.4

29.6

29.8

Fig. 11 Rotation signals at a road joint
0.6High (x10)ABS

Rotation speed s-1

0.5

4. Conclusion

0.4

In this article, we have described an overview of thehigh-resolution rotation sensor integrated hub
bearings and introduced examples of run data with thetest vehicle. It was clarified that sensing response andaccuracy of extracted information can be improved bydetecting rotation speed of the wheels at highresolution. We strive for broad adoption of thistechnology as a means for contributing to improved
vehicle safety including deployment in the control
applications of the electric vehicles.

0.3

High-resolution sensor

0.2
0.1
0
0.1

ABS sensor
0.2

0.3

0.4

0.5

0.6

Time s

0.7

0.8

0.9

1.0

Fig. 10 Slipping on the low-friction surface-56-

*09E_*09 14/08/01 10:38 ページ 6

Hub Bearing with an Integrated High-Resolution Rotation Sensor

References
1) Eiji Funahashi, “History of Hub Bearings andRecent Technology”, NTN TECHNICAL REVIEW
No. 70, pp 52-57 (2002)
2) P.Desbiolles, A.Friz,“Development of High Resolution
Sensor Element MPS40S and Dual Track Magnetic
Encoder for Rotational Speed and Position
Measurement”, NTN TECHNICAL REVIEW No.75
pp36-41 (2007).

3) C. Duret, S. Ueno, “TMR: A New Frontier for
Magnetic Sensing”, NTN TECHNICAL REVIEW No.
80 pp64-71 (2012)4) C.Duret, J.Paul, B.Negulescu, M.Hehn,“TMR: A New
Frontier for Magnetic Sensing”, Proc. of 11th MR
Symposium, Wetzlar, 2011.

5) H. Ito, T. Takahashi, P. Desbiolles, C. Peterschmitt,
S. Ueno, “High Resolution Sensor Bearing with an
Index Signal”, NTN TECHNICAL REVIEW No. 78
pp 76-82 (2010)
6) Katsuhiro Asano, Eiichi Ono, Takaji Umeno,
Masaru Sugai, Yoshitoshi Watanabe, “Estimation of
Tire-Road Friction”, Journal of Society ofAutomotive Engineers of Japan, Vol. 61 No.2 pp50-55 (2007)．
7) Takaji Umeno, “Estimation of Tire-Road Friction by
Tire Rotational Vibration Model”, R&D Review of
Toyota CRDL, Vol. 37 No.3 pp 53-58 (2002).
8) NTN-SNR, Patent No. 4777347

Photo of authors

Kentaro NISHIKAWA

Toru TAKAHASHI

Christophe DURET

Chassis Engineering
Automotive Business HQ

New Product Development
R&D Center

NTN-SNR ROULEMENTS
Research & Innovation
Mechatronics-57-

*10E_*09 14/08/01 10:55 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Press Connect Spline Hub Joint

Takayuki NORIMATSU*
Tsutomu NAGATA*In the automotive industry in the world, low fuel consumption is an
important issue , the lightening to become one of the solution is demandedfrom every part constituting a car. NTN developed the lightening of a hub
bearing and the CVjoint in order to meet these demands.
This report introduces "PCS-hub joint" which is lightening by connecting a
hub bearing and CVjoint by "the press connect method" that is NTN originaltechnology.

2.1 Joint methods of components

1. Introduction

The joint methods of H/B and CVJ are describedusing Fig. 1 in the following:High fuel economy is an important requirement in
the automotive industry today. One of the solutions forachieving high fuel economy is a lightweight design,
which is now required for all vehicle components.
In this article, we introduce the “Press Connect
Spline Hub Joint (PCS-H/J)” which is 12% lighter thanour conventional product and eliminates backlash of
the spline joint by connecting the hub bearing (H/B)
and the constant velocity joint (CVJ). NTN’s
technology for the “Press Connect Method” does not
require any change in the assembly process of theautomotive manufacturers.

a) Conventional structure(third generation H/B+CVJ)
Splines are provided on the outer diameter of the
CVJ outer ring stem and on the inner diameter of H/B,
which are then fit together to transmit torque. The
transmission of torque from CVJ to H/B is conducted
through the contact of the spline teeth. However, close
contact of both components is difficult because of themachining accuracy (difference of pitch of the teeth
between the splines). The resulting gap, as shown inthe middle section of the figure, causes circumferential
backlash. This prevents the torque to be borne by theentire spline, which requires the length of the spline
engagement to be increased.

2. Structure and features
NTN has been developing the “V-Series Hub Joint(V-H/J)” 1) as an integrated unit of the H/B and CVJ
connected with the “Press Connect Method”, whichuses a swaged CVJ outer ring stem (Fig. 1b) toassemble the third generation H/B+CVJ (Fig. 1a)
where conventional H/B and CVJ are fastened
together with a nut. H/B and CVJ are inseparable with
V-H/J design and so the vehicle assembly processneeded to be changed at the automotive
manufacturers. To solve this problem, the newly
developed “PCS-H/J” is designed to fasten the H/B
and CVJ with a bolt which also assisted in reducing
the spline pressure load (Fig. 1c).

b) V-H/J
The CVJ outer ring stem is machined with special
splines (teeth) that are thermally hardened andinserted into the H/B. This causes splines to beformed on the inner diameter of the H/B spindle ringwith which CVJ and H/B are tightly connected. NTN
calls this joint method the "Press Connect" method.
In this method, since the inner diameter of H/B isformed by elastic deformation, the recessed part of the
H/B inner diameter and the raised part of the CVJ
outer ring stem are tightly engaged. The input torque*Chassis Engineering, Automotive Business HQ-58-

*10E_*09 14/08/01 10:55 ページ 2

Press Connect Spline Hub Jointc) PCS-H/J

can be uniformly borne by the entire teeth area
allowing shorter spline engagement length comparedto the conventional structure a).
Since the entire teeth area has to be formed, highpower is required for the machining process including
the use of machining equipment such as a pressmachine for the joint process, which is a challenge.

For PCS-H/J, splines are pre-formed on the innerdiameter of H/B with smaller tooth width than the CVJ
outer ring stem splines (Fig. 2 shows an example ofpre-formed splines). The CVJ stem teeth and thesplined grooves of the H/B are gapped at specified
distances to reduce the power required for theassembly process of the H/B and CVJ that are
fastened together by a bolt. With this design, thevehicle production line at the automotive
manufacturers does not required to be changed for
assembling PCS-H/J.

Structurea) Conventional structurec) PCS-H/J

L2

L3

X

X

X

SwagedX

H/B

Contact condition of teeth(X-X section)b) V-H/J

L1

Gap between teeth

CVJ

W1
W2

Bolt

X

Spline boundary

H/BNo gap between teeth

CVJ

W2
W1 ＞ W2

Gap between teeth

X

Spline boundary

No gap between teeth
W2

W1

W1 ＜ W2

Interference

W1

Gap

Interference

W1: H/B spline arc tooth gap (red), W2: CVJ stem spline arc tooth thickness (blue)
Spline press
fit length

L1 ＞ L2 ＝ L3

Fig. 1 Contact condition of teeth

Hub bearings (H/B)

Constant velocity joint (CVJ)

Hub bearings (H/B)
Pre-splines
Fastening bolt

Stem splines

Press connect joint

Constant velocity joint (CVJ)

Fastening bolt

Fig. 2 Structure of PCS-H/J-59-

Gap H/B

CVJ

W1 ＜ W2

*10E_*09 14/08/01 10:55 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

2.2 Features

With PCS-H/J, input torque is borne by the entire
area of the teeth as opposed to the conventional
structure, therefore, the length of the CVJ stem can be
reduced by approximately 65%. Furthermore, theweight was reduced to 0.94 kg/unit (approx. 12%) by
thinning the inside wall of the hub ring and by using abolt instead of a nut.

PCS-HJ achieved a significant lightweight andeliminated backlash on the splined area thanks to theadoption of the press connect method. Fig. 3 shows
an example of the trial application on a C-segment
vehicle.

PCS-H/J

Conventional structure (third-generation H/B+CVJ)
Thinnerhub ring

100

45

φ80

100

φ80

45

100

Nut：M22×1.5

Mass※1: 4.35 kg

Shorter CVJ stem

35

Bolt：M12×1.25

Mass: 3.88 kg (reduction of 0.94 kg/unit)*1: Mass of H/B+Outboard CVJ outer ring+nut or bolt

Fig. 3 Example of the application of PCS-H/J

W2(Interference of press connect: W2−W1)

3. Functional test
W1

W1 : H/B spline arc tooth thickness (red)

3.1 Spline press-fit load

W2 : CVJ stem spline arc tooth thickness (blue)

With PCS-H/J, the CVJ outer ring stem isassembled to the inner diameter of H/B by use of abolt. Fig. 5 shows the measurement results of the
spline press-fit load using components with the upperand lower limits of the interference as shown in Fig. 4.
Fig. 5 shows that the spline press-fit load required
for the upper and lower limits of interference is below
the lower limit of the established axial force of the bolt
which means that the H/B and CVJ can be assembled
in all specified ranges of interference.

Interference

Fig. 4 Interference

Spline press-fit load kN

Interference standard
Lower limit ofbolt axial force
Upper
limit

Lower
limit
Interference mm

Fig. 5 Spline press fit test-60-

*10E_*09 14/08/01 10:55 ページ 4

Press Connect Spline Hub Joint

3.2 Strength evaluation of the splined area

3.3 Disassembly/reassembly test

Static torsional and bi-directional torsional fatigue
strength tests were conducted against the shortened
splined area of PCS-H/J and had verified the strength
was equal or better than the conventional unit.

Service or replacement of the PCS-H/J components
was studied. Disassembly of the H/B and CVJ was
conducted and the spline press-fit load was evaluatedwhen they were reassembled. The spline strength ofthe disassembly and reassembly unit was verified to
be OK.

3.2.1 Static torsional strength test
Static torsional strength test was conducted forbreaking torque when torsional torque is applied in
one direction. Fig. 6 shows the test results.
Although the splined area is short, PCS-H/J has
equal or better strength than the conventional units.

3.3.1 Measurement of spline press-fit load
during reassembly after disassembly
In this test, spline press-fit load was measured withthe combination of both new and reused H/Bs andCVJs, as shown in Table 1, and verified to see ifinterference still remained in the splined area of thereused components. Fig. 8 shows the test results.

Breaking torque kN・m

Table 1 Sample combination
CVJ

H/B

PCS-H/J

Fig. 6 Static torsional strength test

3.2.2 Bi-directional torsional fatigue strength test
A bi-directional torsional fatigue strength test was
conducted where alternating torsional torque isrepeatedly applied on the splined area and records thecycle count until the unit breaks. Fig. 7 shows the test
results.
Both high torque and low torque were applied
repetitively for the evaluation with the assumption thathigh torque and low torque is what is normally seen on
vehicles. PCS-H/J demonstrated to have sufficientfatigue strength over the guaranteed life values of theconventional products even if the splined length isshorter.

Reused5 times

New

(1)

(2)

Reused5 times(3)(4)(1) H/B New-CVJ New(2) H/B New-CVJ Reused 5 times(3) CVJ New-H/B Reused 5 times(4) H/B Reused 5 timesCVJ Reused 5 times

(1)

(2)

(3)

(4)

Fig. 8 Re-assembly load test

When only the CVJ was the reused component(Fig. 8 (2)), the press-fit load was almost the same asin the case of new components when comparing thespline pressure load of new components (Fig. 8 (1)).
When only the H/B was the reused component (Fig. 8(3)), the press-fit load was about half of the newcomponent. The press-fit load decreased because theinterference had been reduced as the H/B innerdiameter was formed by the CVJ splines when theoriginal CVJ was pressed in. When both H/B and CVJ
were the reused components (Fig. 8 (4)), the press-fit
load was approximately 1/3 of the new components.
Based on the above data it was revealed that when
the H/Bs were reused, the press-fit load is reduced
compared to the new components, however, the load
is not zero and some interference still remained in the

PCS-H/J
Load torque kN・m

New

Spline press-fit load kN

Conventional unit

Numberof reuse

Guaranteed value ofthe conventional units

Number of cycles

Fig. 7 Bi-directional torsional fatigue strength test-61-

*10E_*09 14/08/01 10:55 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

Interference

splined area. In addition, the reused components wereverified to be fit without any backlash being observedin the joint.

Roundness 0.8μm

Roundness 1.1μm

Roundness 1.4μm

Roundness 1.6μm

After press connect

In order to verify the strength of the splined area ofthe disassembled/reassembled PCS-H/J, samples
which have undergone disassembly/reassembly 5
times were placed on the bi-directional torsional
fatigue strength test of the above section 3.2.2. Fig. 9shows the results.
They were evaluated with high torque and low
torque, similar to 3.2.2. The disassembled andreassembled unit has the equivalent level of strength
as the first assembled (new) unit.
Based on the above data it was observed that thereuse of H/B and CVJ exhibits no problem forassembly or strength.

Fig. 10 Circularity measurement of raceway surface

Newly assembled (Table 1 (1))
Reassembled 5 times (Table 1 (4))

Guaranteed value ofthe conventional units

Number of rotations rev

Load torque kN・m

Lower limit

Before press connect

3.3.2 Bi-directional torsional fatigue strength testof disassembled and reassembled units

Upper limit

Number of cycles

Fig. 9 Bi-directional torsional fatigue strength test

3.4 Verification of impact to H/B
Inspection of the raceway condition after bearings
life testing was conducted to verify the impact of press
connect on H/B.

Conventionalunit

PCS-H/J

Fig. 11 Turning (0.8G) durability test

3.4.1 Measurement of roundness of raceway

Axial force residual ratio %Roundness of raceway of the inner ring before andafter press-fitting splines with upper and lower limits ofinterference (Fig. 4) was measured. Fig. 10 shows theresults.
The roundness measurement value after pressfitting the splines increased by approximately 0.5 μm.
No impact such as polygon strain was observed. Also,no difference between press connect interference wasobserved.

3.4.2 Turning life test with 0.8G
A load equivalent to 0.8G which simulates
aggressive turning was applied to the PCS-H/J
bearings so life testing could be conducted. Fig. 11
shows the results.
The bearing life was equivalent to the conventional
products and no difference was exhibited.

Conventionalunit

PCS-H/J

Fig. 12 Axial force residual ratio-62-

*10E_*09 14/08/01 10:55 ページ 6

Press Connect Spline Hub Joint

3.5 Measurement of bolt axial force
Since the PCS-H/J uses a bolt for fastening insteadof the conventional use of nuts (Fig. 1), loosening of
the bolt during use in the field is a concern.
Bi-directional torsional torque was applied on the
splined area of PCS-H/J for predetermined number oftimes and bolt axial force was measured before andafter the test to verify if the bolt had loosened. Fig. 12shows the results.
No loosening was observed after the test and theresidual ratio of the bolt axial force was 86% or morefor PCS-H/J as opposed to the conventional nut
fastening of approx. 80%. This is considered to beadequate for operational usages.

4. Summary
We have developed a lightweight PCS-H/J that has
no backlash in the splined area by forming pre-splines
on the inner surface of the H/B with smaller tooth
width than the splines of the CVJ stem, and uses abolt for connecting the H/B and CVJ.
PCS-H/J can be assembled within the existingassembly process in the vehicle production lines ofautomobile manufacturers. We have pursued weight
reduction with PCS-H/J which is desired for themodern vehicles and believe it will significantly
contribute to improved fuel economy.
We plan to expand this PCS-H/J into a family of
product lines and continue developing modular units.

References
1) Mitsuru UMEKIDA, Yuuichi ASANO, “V-series Hub
Joint”, NTN TECHNICAL REVIEW No. 77 p67-72,
2009

Photo of authors

Takayuki NORIMATSU

Tsutomu NAGATA

Chassis Engineering
Automotive Business HQ

Chassis Engineering
Automotive Business HQ-63-

*12E_*09 14/08/01 11:02 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Next-generation High Efficiency Fixed TypeConstant Velocity Joint "CFJ"Teruaki FUJIO*

The demand of low fuel consumption to the car from consideration ofenvironment has increased in recent year. In NTN, in order to realize
high efficiency of drive shaft, we developed a fixed constant velocity
joint which reduced torque loss by original design. This paper
introduces the feature and performance of the CFJ.

In this article, we introduce features andperformance of "CFJ", which reduced torque loss byhalf against EBJ by adopting NTN’s original design toreduce internal friction loss of the fixed-type CVJ.

1. Introduction
The common challenge of the world automotive
industry for achieving “high fuel economy” is to reduce
the loss of torque from the components that comprisethe powertrain.
The power from the engine is transmitted to the tires
through the transmission, differential gear and drive
shaft. This drive shaft is called constant velocity joint(CVJ) because it smoothly transmits torque at aconstant speed even when the input differential gear
shaft and output shaft on the tire side rotate atdifferent angles. CVJ is usually configured with afixed-type CVJ, which enables a greater joint angle but
is not capable of axial sliding, and a sliding-type CVJ,
which enables only a small joint angle but is capableof axial sliding, connected by a shaft (Fig. 1). As
shown in Table 1, the torque loss of the CVJ has beensmall among the powertrain components; however,further improvement of efficiency is now required foroverall high fuel economy.
Fig. 2 shows the history of weight reduction andefficiency improvements of fixed-type constant velocity
joints at NTN. NTN has been working to enhance thefixed-type CVJ and was among the first to develop thelightweight, compact, and highly efficient E-Series 3)
while contributing to the improvement of the vehicle
fuel economy. On the other hand, the torque loss ofthe sliding-type CVJ is small compared with the fixedtype CVJ (about 0.2%) as the tripod type withintegrated rolling bearings has been in the mainstream.

Tire
Differential gear

Drive shaft

Input
shaft

Outputshaft

Sliding-type CVJ
Shaft
Fixed-type CVJ

Fig. 1 Layout of drive shaft 1)

Table 1 Torque loss of drive-train 2)*Drivetrain Engineering, Automotive Business HQ-64-

Component

Torque loss rate

Transmission

1〜5%

Differential gear

1〜5%

Drive shaft (CVJ)1%*12E_*09 14/08/01 11:02 ページ 2

Next-generation High Efficiency Fixed Type Constant Velocity Joint "CFJ"

Conventional
unit!

10％

20％
Developed unit

90％

(6 balls)

80％

Conventional
unit@

Weight ratio

(%)
100

Next generationhigh efficiency
CVJ

High efficiency:
E-Series(6 balls)(8 balls)

99.5

︵
︶

︵
︶

99.0

Downsizing by highstrength shaft andhigh-performancegrease1990’s2000’s2011

Era

Transmission efficiency

100％

2012

Fig. 2 Transition of the weight saving and increase in efficiency of NTN’s fixed constant velocity joint.

2. Structure and features of CFJ
2.1 Structure of CFJ
CFJ consists of an inner ring/outer ring equippedwith a ball raceway groove (track), eight balls thattransmit running torque, and a cage that holds theballs (Fig. 3).

Outer ring

Outer ring

Inner ring

Fig. 4 Configuration of track form
Ball

Outer track

Inner track

Cage
Inner ring

Fig. 3 Configuration of CFJ

Ball

2.2 Features of CFJ
For reduction of torque loss, we have adopted ourunique structure to reduce friction of contact points in
CVJ.
Feature 1: Circular arc tracks of the inner andouter rings are inclined to the axial
direction and the adjacent tracks arearranged in mirrored symmetry (Fig. 4).
Feature 2: The slopes of the inner and outer tracks
are crossed with each other and a ballis placed in the intersection (Fig. 5).

Fig. 5 Layout of track and ball-65-

*12E_*09 14/08/01 11:02 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

opening side (F1) and the reverse side (F2). As F1and F2 have the same magnitude and are applied in
the opposite directions, they are cancelled out and thecage axial displacement is limited. As a result, thespherical contact force generated between the cageand outer ring, as well as the cage and inner ring canbe significantly reduced.

3. Mechanism of reduction of torque loss
Torque loss of a CVJ is caused by the energy loss
due to the friction of contact points.
Fig. 6 shows the contact conditions of the current
EBJ. The tracks of the inner and outer rings areformed on the circular arc and their centers are
displaced on the opposite directions from the joint
center. The ball that transmits power is held by the
tracks of inner/outer rings which form a wedge angle(τ) and is supported by the cage. When torque isapplied to the joint, the tracks of inner and outer rings
and the ball make contact and the ball is given force(F) to push the cage because of the wedge angle.
Then, another load that balances with this force isgenerated between the cage outer spherical surface
and the outer ring inner spherical surface, as well as
the cage inner spherical surface and the inner ringouter spherical surface.
In the development of CFJ, we focused on the
reduction of friction resistance of sliding motion on thespherical surface (Fig. 6 (4) and (5)).
Fig. 7 shows the inner force that is generated. As
mentioned in Section 2, adjacent circular arc tracks
are mirrored-symmetrically placed in the CFJ layout.
The direction of the force that the balls push the cage(F) can be alternately distributed to the outer ring

F1
F2

Outer ring
Track onthe load side

Outer ring
Track onthe load side

Fig. 7 Contact force of CFJ

4. Verification by analysis
Fig. 8 is a comparison of contact force of sphericalparts between CFJ and EBJ which is obtained bystructural analysis. It shows that the contact force of
spherical parts of CFJ is significantly low at any joint
angles.
We could verify that the mechanism of the torqueloss reduction by our unique structure is working as
designed from the analysis result as well.
We also obtained similar results for the force
between the cage and inner ring.

Ball

Outer ring

(1)
(5)

Cage

τ
Inner ring

(4)

Center of
inner ring track[Contact points]

(2)

CFJ
EBJ (E-Series)

Contact force of spherical part(Cage outer spherical surface andouter ring inner spherical surface)(3)

Center ofouter ring track
Center of joint(1) Outer ring track and ball(2) Inner ring track and ball(3) Cage and ball(4) Cage inner spherical surface and inner ring outer spherical surface(5) Cage outer spherical surface and outer ring inner spherical surface

0(For simplification, contact points between track and ball are indicated on the bottom of track)

Fig. 6 Composition of EBJ

2

4

6

8

Joint angle

10

°

12

14

Fig. 8 Analysis result of contact force of spherical part-66-

*12E_*09 14/08/01 11:02 ページ 4

Next-generation High Efficiency Fixed Type Constant Velocity Joint "CFJ"

5.2 Increase of temperature

5. Performance characteristics

Fig. 10 shows a comparison of increase of
temperature between CFJ and EBJ.
CFJ is able to reduce not only torque loss but also
self-heating, which reduces the joint surface
temperature by around 50% compared to EBJ. This
allows expanded application of operating range in highrotational speeds.

5.1 Transmission efficiency (torque loss rate)
Fig. 9 shows a comparison of torque loss rates
among CFJ, EBJ, and BJ which is used worldwide asconventional joint.
It shows that the current EBJ has lower torque loss
rate and higher efficiency than BJ. We could verify that
CFJ’s torque loss rate is significantly lower than EBJ
at the joint angles of more than 6 degrees.
In general, torque loss rate deteriorates as the joint
angle increases because of the increased relative
motion of the internal components; however, CFJ is
not very affected by the joint angles, thus maintaining
a low torque loss. Therefore, CFJ is able to provide ahigher loss reduction effect when installed on thevehicles where the joint installation angles become
relatively large such as SUVs having a greater vehicle
height.

6. Summary
We have developed the “CFJ” which is a nextgeneration high efficient fixed type CVJ that reduces
torque loss rate by 50% compared to the current ESeries at the joint operating angle range of over 6
degrees. It achieved the highest level of efficiency andheating reduction in the industry by reducing
significant internal friction loss using our uniquedesign.
NTN would like to contribute to further global
environmental protection by expanding this nextgeneration fixed type CVJ with its superiortransmission efficiency in the global market in
response to the ever-increasing demand for improved
fuel economy.

2.0

Torque loss rate %CFJ
EBJ (E-Series)

1.5

BJ (Convention joint)

1.0

References

0.5

0

0

2

4

6

8

Joint angle

10121) Manabu HOSHINO, Masashi FUNAHASHI: Fixed
Constant Velocity Joint with a Super High Operating
Angle of 54 Degrees (TUJ), NTN TECHNICALREVIEW No. 75 (2007), 16-19

14°2) Society of Automotive Engineers of Japan:
Automotive Engineer’s Handbook, Toshoshuppan,
Tokyo (1979), 3-35.

Fig. 9 Comparison of torque loss rate3) Keisuke SONE, Kazuhiko HOZUMI: High Efficiency
Compact Fixed Constant Velocity Universal Joints(EBJ), NTN TECHNICAL REVIEW No. 66 (1997),
28-31

Joint temperature ˚C

120
100
80

Reduction byapprox. 50%

60

Photo of author

40
20
0

CFJ

EBJ (E-Series)

Fig. 10 Comparison of joint surface temperature

Teruaki FUJIODrivetrain Engineering
Automotive Business HQ-67-

*13E_*09 14/08/01 11:07 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]Development of Low Torque Seal Ring
for Automotive Transmission
Kouzou KAKEHI*
Takumi KONDOU*
Takuya ISHII**
Masato YOSHINO***
Further lower level of torque and oil leak is required of seal rings used in
AT and CVT as fuel economy standards of cars is higher in the future.
Therefore, NTN developed “Low Torque Seal Ring”, which is made of PEEK
resin and equipped with specially-shaped lubrication grooves by injection
molding, keeps low torque (60% reduction) and low wear properties incomparison with NTN conventional products and keep its oil leakage characteristic equal to conventional
one. This article introduces the characteristic and performance of “Low Torque Seal Ring”.

energy. Low oil leakage allows for high efficiency andsmaller size of the oil pump, resulting in energy
savings. In order to maintain low torque and low oilleakage long-term, and extend the life of thecomponents, it is necessary to minimize wear of theseal rings and components on which the seal rings
slide.
Seal ring materials are mainly polytetrafluoroethylene (PTFE) resin and PEEK resin. PTFE resin
seal rings are made by machining compression
molded material, while PEEK resin seal rings are
made by injection molding. The seal rings made of
PEEK resin with complex abutment geometry frominjection molding exhibit lower oil leakage than theseal rings made of PTFE resin with simple and straightabutment geometry.

1. Introduction
Automatic transmissions (AT) and continuously
variable transmissions (CVT) use 4 to 10 resin sealrings of 15 to 60 mm outer diameter. As stricter fueleconomy standards are expected to be introduced forvehicles, seal rings are required to have even lower
torque and lower oil leakage characteristics.
In this article, NTN is introducing the “Low TorqueSeal Ring” with high functionality by using speciallydesigned lubrication grooves to the seal rings made ofpolyether ether ketone (PEEK) resin 1), 2).

2. Roles and challenges of seal rings
2.1 Roles of seal rings
Seal rings are moving seals which seal oil in a
sliding action. Seal rings are installed between
relatively moving shafts and housings in the hydrauliccircuit of automatic transmissions (AT) and CVTs, as
shown in Fig. 1. They seal oil and maintain pressure
within the hydraulic circuit while being pressed to boththe housing and shaft by hydraulic pressure in a
sliding motion.
The required performance for seal rings is low
torque, low oil leakage and low friction. Low torqueimproves the transmission efficiency and saves

Seal rings

Fig. 1 Structure of automatic transmissions***Engineering Dept, NTN Engineering Plastics Corp.
***Composite Material Engineering Dept, Composite Material Products Division***Advanced Technology R&D Center-68-

*13E_*09 14/08/01 11:07 ページ 2

Development of Low Torque Seal Ring for Automotive Transmission

2.2 Challenges of seal rings

3. Features of V-shape groove seal ring

In general, the cross-section shape of a seal ring iseither rectangular or trapezoidal. The respective
operations are shown in Fig. 2.
The rectangular seal ring is pressed by hydraulicpressure to the inner surface of the housing and theside wall of the shaft groove. The contact area
between the seal ring and the shaft groove side wall issmaller than the contact area between the seal ringand the housing inner surface. Therefore, when the
housing or the shaft rotates, the seal ring side and theside wall of the shaft groove slide due to their smaller
sliding resistance. Since the contact area with theshaft groove side wall is large, the leakage is small but
the torque is high.
The trapezoidal seal ring also slides on the shaft
groove side wall similar to the rectangular seal ring.
But since there is a gap between the seal ring sideand the shaft groove side wall, and hydraulic pressureis also applied from that side, the torque is reduced
compared to the rectangular seal ring. However, thesmall contact area with the shaft groove side wallcauses higher and unstable oil leakage. In addition,wear of the seal rings makes the contact area larger,
resulting in gradually increased torque, closer to thevalue of the rectangular seal rings. Therefore,
rectangular seal rings, which have more stable torqueand oil leakage, are used more often.
In order to solve these issues, NTN has developed
a seal ring that achieves both high sealing and low
torque properties by providing a V-shape groove on
the side of the seal ring (referred to hereafter as Vgroove seal ring).

Fig. 3 shows the structure of the V-groove seal ring.
Providing the V-shape lubrication groove on the sidewall of the rectangular seal ring makes the contact
area with the shaft groove side wall smaller than therectangular seal ring, and optimizing the V-shapelubrication groove geometry enables lubrication on thesliding surface, achieving low-torque and low-wear
properties.
The material of this developed product is Bearee
PK5301, which is based on PEEK resin with a special
filler blended in.
Table 1 shows a comparison of the developed Vgroove seal ring with the rectangular and trapezoidal
seal rings. The V-groove seal ring has a good
performance balance and the following advantages
against the rectangular seal ring (NTN conventional
product).

(Features)(1) 60% lower torque(2) 1/10 wear(3) Equivalent low oil leakage

Abutment
V-shape lubrication groove

Trapezoidal seal ring

Housing

Housing

Seal ring

Shaft

Seal ring

Table 1 Comparison of features various seal rings

Shaft grooveside wall

Rectangular seal ring

Shaft grooveside wall
Lubricating oil

Lubricating oil

Fig. 3 Groove geometry of low torque seal ring

Item

Rectangularseal ring

Trapezoidal
seal ringV-grooveseal ring

Cross-section
shape

RectangularTrapezoidal

RectangularSide surface
lubrication groove

None

None

V-shapeContact with shaft
groove side wall

Full surface
contact

Line contact

Full contact otherthan groove area

Torque△◎〜△○Oil leakage

◎

△

◎

Wear properties
Performancebalance

○

△

◎

○

△

◎

◎：Excellent○：GoodPartialcross-section
diagram

Shaft

Fig. 2 Operating state of seal ring-69-△：Fair*13E_*09 14/08/01 11:07 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

Table 3 Material of test pieces

4. Performance of V-groove seal ring

Material

Test piece

4.1 Test seal rings

Seal ring

PEEK resin (Bearee PK5301)

NTN has tested three types of seal rings, as shownin Table 2, to verify the performance of the V-shapelubrication groove. Since the lubrication groove
section of the V-groove seal ring does not makecontact with the shaft groove side wall, the torque islower than the rectangular seal rings, similar to thetrapezoidal seal rings. In addition, the V-shapelubrication groove improves torque and anti-wear
properties compared with the other lubrication groove
geometries. Therefore, in order to verify the differenceof the groove geometries, NTN also evaluated squaregroove seal rings with square lubrication grooves withthe same contact area to the shaft groove side wall as
the V-groove seal ring.

Housing

S45C

Shaft

S45C

4.3 Torque4.3.1 Dependency on hydraulic pressure
Torque (for two seal rings) was evaluated under thevarying hydraulic pressure test conditions shown in
Table 4. Fig. 5 and 6 show the relationship between
hydraulic pressure and torque.
The torque of the V-groove seal rings wasapproximately 60 to 70% lower than that of therectangular seal rings in any rotation speed andhydraulic pressure, and dependency on the hydraulicpressure was also relatively low. With V-groove sealrings, the higher the hydraulic pressure, the more oilenters onto the sliding surface from the lubrication
groove Therefore, the torque dependency on hydraulicpressure is considered to be smaller.
From the groove geometry standpoint, the V-groove
seal rings showed about 20% lower torque than that of
the square-groove seal rings.

Table 2 Test seal rings
Test seal ringGeometry of side surface lubrication groove

V-groove seal ring

V-shape

Square-groove
seal ring

Squareshape

Rectangularseal ring−(No groove)

Table 4 Test condition

Seal ring size: OD 50 mm, Thickness: 1.6 mm, Width: 1.5 mm

Item

Description

Hydraulic pressure

0.5〜2.0MPa

Rotation speed

4,000min-1 or 7,000min-1

Fig. 4 shows an overview of the test machine which
evaluates torque, oil leak and wear.
Pressure was applied while supplying oil between
two seal rings mounted on the shaft grooves androtating the housing. Torque and oil leakage were
evaluated during this operation. Table 3 shows thematerials of the seal rings, housing and shaft.

Lubricating oil

ATF (temperature: 110°C)

Torque (for two seal rings) N・m

4.2 Test machine and conditions

Seal rings
Housing (rotating)

Motor

Torque
meter

Shaft

Hydraulicpressureunit(oil circulation)

1.5

Rectangular seal ring
Square-groove seal ringV-groove seal ring

1.0

0.5

0

0

0.5

1.0

1.5

2.0

Hydraulic pressure MPa

Fig. 5 Torque versus oil pressure (4,000min-1)

Oil leakage

Fig. 4 Test machine

-70-

2.5

*13E_*09 14/08/01 11:07 ページ 4

Development of Low Torque Seal Ring for Automotive Transmission

4.4 Oil leakage

Torque (for two seal rings) N・m

1.5

Rectangular seal ring
Square-groove seal ringV-groove seal ring

1.0

V-groove seal rings and rectangular seal rings were
evaluated for oil leakage under the varying
temperature test conditions shown in Table 6. Fig. 8
shows the relationship between oil temperature and oilleakage. The decrease of lubricating oil viscosity alongwith the rise of temperature causes oil leakage toincrease. The oil leakage of the V-groove seal rings
was very small compared to that of the rectangularseal rings.

0.5

0

0

0.5

1.0

1.5

Hydraulic pressure MPa

2.0

2.5

Fig. 6 Torque versus oil pressure (7,000min-1)

Table 6 Test condition
Description

Item

4.3.2 Dependency on oil temperature

Oil leakage amount ml/min

Torque (for two seal rings) was evaluated under thevarying oil temperature test conditions shown in Table
5. Fig. 7 shows the relationship between oil
temperature and torque. The torque of the V-groove
seal rings is lower than that of the rectangular sealrings by 60% at any temperature. From the groove
geometry standpoint, the V-groove seal rings showedabout 20% lower torque than that of the squaregroove seal rings.

Table 5 Test condition

Torque (for two seal rings) N・m

Hydraulic pressure

2.0MPaRotation speed

4,000min-1

Lubricating oil

ATF (temperature: 30-150°C)

Lubricating oil

ATF (temperature: 30-150°C)

30
V-groove seal ring25

Rectangular seal ring

20
15
10
5
0

0

40

160

Wear of seal rings was measured under the highpressure, high speed and high temperature test
conditions shown in Table 7. Fig. 9 shows the wear of
the side walls of the seal rings.
Wear on the low-torque V-groove and squaregroove seal rings with lubrication grooves was smaller
than that of the rectangular seal rings.
Also, the V-groove seal rings showed lower wear
between two different groove geometries at 1/10 thatof the rectangular seal rings.
The addition of a lubrication groove improves the
lubrication of the sliding surface and reduces wear. It
can be seen that the groove geometry for lower torqueresults in lower wear as well.

0.5

80

120

4.5 Wear properties

1.0

40

80

Fig. 8 Oil leakage versus oil temperature

Rectangular seal ring
Square-groove seal ringV-groove seal ring0

4,000min-1

Temperature ˚C

1.50

2.0MPaRotation speed

Description

Item

2.0

Hydraulic pressure

120

160

Oil temperature ˚CFig. 7 Torque versus oil temperature-71-

*13E_*09 14/08/01 11:13 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

Table 7 Test condition

Table 8 Analysis condition

Item

Description

Hydraulic pressure

3.0MPaSliding area

Rotation speed

7,000min-1

Oil film thickness

Lubricating oil

ATF (temperature: 150°C)

Shaft rotation speed

4,000min-1

1hHydraulic pressure

2MPa

Oil temperature

20˚C

Time

80

Amount of wear µm

Description

Item

Only on shaft groove side wall
5µm

V-groove seal ring

60
40

Rotational direction
20
0

V-grooveseal ring

Square-groove
seal ring

Square-groove seal ring

Rectangulaseal ring

Fig. 9 Wear of test seal rings
Rotational direction
Low ←4.6 Fluid analysis

Pressure -4000

The test confirmed that the developed V-shapegroove of the V-groove seal rings was effective in
reducing torque. This result is due to the reduced
contact area with the shaft groove side wall andimproved lubrication of the sliding surface by additionof the lubrication grooves. The torque difference due
to groove geometry was assumed to be caused bydifferent lubrication conditions. In order to prove thistheory, NTN extracted and modeled the fluid pattern of
one lubrication groove for the V-groove seal ring andsquare-groove seal ring and conducted a fluid
analysis.
Table 8 shows the analysis condition and Fig. 10shows the analysis result. The ends of the V-shapegroove of the V-groove seal ring show high pressure
due to hydrodynamic effects. This force is in the
opposite direction of the hydraulic pressure applied tothe seal ring which pushes on the shaft groove side
wall, contributing to the reduction of pressure on thecontact surface. In addition, it facilitates oil supply fromthe ends of the lubrication groove to the sliding
surface, which contributes to lower torque. On theother hand, high pressure similar to the V-shapegroove is not shown at the ends of the square-groove
seal ring. Thus the different lubrication groove
geometries resulted in the experienced torquedifferences.

Pa→ High6.72e+51.35e+6

2.02e+62.70e+6

Fig. 10 Static pressure distributions of sliding width slide
surfaces-72-*13E_*09 14/08/01 11:07 ページ 6

Development of Low Torque Seal Ring for Automotive Transmission

5. Summary
NTN has developed V-groove seal rings which
reduced oil leakage and significantly improved thetorque and wear properties of the rectangular sealrings. They are currently being evaluated byautomotive and transmission manufacturers and NTN
believes they will contribute to low fuel consumption.
NTN strives to continue with development of lower
torque and longer life products by applying thislubrication groove design technology to bearings andother units used in fluid lubrication.

References
1) Yutaka HAYASHI, Jiro MATSUMOTO, Noboru
UMEMOTO: Introduction of Engineering Plastic
Seals, NTN TECHNICAL REVIEW No. 65 (1996)
54.
2) Yoshio OKI, Takumi HAYASHI, Takuya ISHII:
Characteristics and Applications of Super
Engineering Plastics, Journal of Japanese Societyof Tribologists 49 (7), (2004) 31

Photo of authors

Kouzou KAKEHI

Takumi KONDOU

Takuya ISHII

Masato YOSHINOEngineering Dept.
NTN Engineering Plastics Corp.

Engineering Dept.
NTN Engineering Plastics Corp.

Composite Material Engineering Dept.
Composite Material Products Division

Advanced Technology
R&D Center-73-

*14E_*09 14/08/01 11:17 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Product ]The Introduction of Sintered New Products
for Automobile
Tomonori YAMASHITA*
Tomokazu SONOZAKI**
The sintered machine part becomes high-strength, and the adoption of
the automobile use increases recently. A sintered part had characteristicthat it could produce the parts of a complicated shape with high productivity,
but securing of strength of the part was difficult to have the porous comparedto cast alloy. Therefore, NTN reviewed the material and process of forming
and sintering at first, and developed that raises the strength of sinteredmaterials by doing it densely. In addition, NTN introduces about the sintered materials for the special
environment as corrosion environment or keeping away lubricant.

the representative methods.
Powder forging, shown in Fig. 1, is a method toincrease the density of sintered compacts by hot forging
with the intent of completely eliminating the internalpores. Copper infiltration is a method wherein theinternal pores of sintered compacts are filled withmelted copper. Warm compacting is a method toincrease the density by heating the powder and moldand using a special lubricant. These conventional highdensity methods listed require special processes whichlead to higher costs.
Therefore, we aimed for obtaining high-density
sintered compacts with a similar method as the ordinary
sintering process of the past. This was done by
optimizing manufacturing conditions, such as material
composition, forming conditions, heat treatment, etc.

1. Introduction
The use of sintered machine parts in vehicles hasgrown in recent years. Domestic production grew by
20% from approximately 70,000 tons in 1998 to 84,000
tons in 2012. Total weight of sintered parts in a vehicle
also grew by 50% from approximately 6 kg in 1998 to
approximately 9 kg in 20111). In this article, we would
like to introduce high-density and high-strength sinteredmachine parts which can be used in a variety of vehicle
applications along with sintered bearings, which can beused in corrosive or poorly lubricated environments.

2. High-density/high-strength sinteredmaterials
Sintered materials are used in automotive
applications in engine and drivetrain components wherehigh strength is required in addition to high reliability
and quality are required. In this section, we will describe
the development of high-density and high-strength
sintered materials.

High

Cost

Powderforging

2.1 High-strength sintered materials byhigh-density process

Copper infiltration
Warm

Sintered compacts, which contain internal pores,
have inferior mechanical properties compared to thematerials made by the melting and forging cast process.
Therefore, various methods for strengthening sinteredmaterials by increasing density are being proposed.
Fig. 1 shows the relation between density and cost of

Low

Conventional
compacting
manufacturing
process
One-time forming
One-time sintering

Low

Density

Developed
material

High

Fig. 1 Relationship between density and costs**Engineering Dept. NTN Powder Metal Corporation**Composite Material Engineering Dept, Composite Material Products Division-74-

*14E_*09 14/08/01 11:17 ページ 2

The Introduction of Sintered New Products for Automobile

Radial crushing strength MPa

and high-density/high-strength sintered materials weredeveloped.

2.2 Mechanical properties
Fig. 2 shows the relation between the density ofvarious materials, including the developed material,
and Young’s modulus. The density of the developed
product is 7.5 g/cm3. This achieves more than 95%, intrue density ratio, of the density of iron, which is 7.85
g/cm3. Young’s modulus is 180 - 190 GPa, which isequivalent to high carbon chromium bearing steel,
SUJ2. Adding cold working further increases bothdensity and Young’s modulus.
Fig. 3 shows the radial crushing strength. Radial
crushing strength is obtained by the following
equation, applying compressive stress on the annular
test piece, as shown in Fig. 4. The developed material
showed a radial crushing strength of 2,000 MPa or
more, compared to 1,300 MPa of the conventional
high-density materials.

2500
2000
1500
1000

Developed
material+
cold working

Developed
material

Conventional
high-density
material

Fig. 3 Radial crushing strength

Load

K : Radial crushing strength (MPa)Maximum load when the test piece is destroyed (N)F（D−e） FL :: Lengthof hollow circular cylinder (mm)K=
D : Outer diameter of hollow circular cylinder (mm)L・e2
e : Wall thickness of hollow circular cylinder (mm)Fig. 4 Method for radial crushing strength test

8.0

600

Fatigue limit

7.8

500

7.6

400

7.4

300

7.2

200

7.0

100

6.8

Conventional Developed Developedmaterial
high-density material+material
cold working

SUJ2

0

Fig. 5 Density and fatigue limit

Fixed side
Stress ratio 0.1

SUJ2

Developed
material

220
200

Conventional
high-density
material

180
160

Developed
material+
cold working

140
120

7.0

7.2

7.4

7.6

Density g/cm3

7.8

Load cell
Annular test piece

Repetitive load

Young's modulus GPa

240

8.0

Movable side

Fig. 2 Relationship between density and Young’s modulus

Fig. 6 Method for fatigue limit-75-

Fatigue limit MPa

Radial crushing strength g/cm3

Fig. 5 shows the fatigue limit of the developedmaterial obtained from the ring compression fatigue
test. The ring compression fatigue strength is obtained
from a test of finding the fatigue limit by repeating aload of stress ratio 0.1 (the ratio between the maximum
and minimum compression load is 10:1) on the annular
test piece, as shown in Fig. 6. The result of the ring
compression fatigue test is, in general, experimentally
known to be around 0.8 times the fatigue strength
obtained by the ordinary bending fatigue test.
The fatigue limit of the developed material is 300
MPa, about twice the fatigue limit of the conventional
high-density material of 150 MPa.
Therefore, the developed material can be used inareas with a high repetitive load, where theconventional sintered parts could not be used.

*14E_*09 14/08/01 11:17 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

The test results showed that the developed material
was equivalent in corrosion resistance against organicacid and superior in corrosion resistance against sulfur.

3. Corrosion resistant sintered bearings
In general, bronze series sintered bearings are used
for motor-based fuel pumps of automobiles. Corrosion
by sulfur and organic acid contained in the fuel has
recently been observed in the conventional bronze
series sintered bearings due to the use of inferior quality
fuel used mainly in developing countries. To cope withthis issue, cupronickel-based sintered bearings withsuperior corrosion resistance are used. However, thismaterial is very expensive as it contains nickel.
This led to the development of the less costly
sintered bearings, adopting aluminum bronze as thematerial, which has equivalent or better corrosion
resistance as the cupronickel material.

3.3 Friction/wear properties
Table 2 shows the friction coefficient and specificwear amount.

<Test conditions>(1) Friction test
・Load: 35N・Rotation speed: 3,160 min-1
・Test piece size: ID 5 × OD 10 × Width 7 mm
・Opposite material: SUS304
・Test time: 30 min.
(2) Wear test
・Load: 64N・Rotation speed: 3,000min-1
・Test piece size: ID 6 × OD 12 × Width 6 mm
・Opposite material: SUS304
・Test time: 500 hours
The friction coefficient of the developed material
was equivalent to that of the cupronickel material. On
the other hand, the specific wear amount of thedeveloped material was about 1/10 that of the
cupronickel material.
This shows that the developed material has a more
superior anti-wear property than the cupronickel
material, thus making it a suitable replacement for thelatter.

3.1 Features and challenges of aluminum bronze
Although aluminum bronze exhibits superior
corrosion resistance, it was considered inadequate forthe sintering process of powder metallurgy because of
surface oxidation under high temperature. It also has ahigher friction coefficient than the ordinary bronze
series material.
To solve this issue, the material composition andmanufacturing method were optimized, especially thesintering method to improve its properties. This achieved
aluminum bronze with corrosion resistance equivalent toor better than that of the cupronickel material. In addition,
friction and anti-wear properties were significantly
improved by optimizing the blending of graphite.

Table 2 Friction coefficient and specific wear rate
Developed material Cupronickel material

3.2 Corrosion resistance
Friction coefficientSpecific wear rate m3/(N・m)

Table 1 shows the test results of corrosion resistanceagainst organic acid and sulfur. Cupronickel material
was also tested for comparison purposes.

0.14

0.14

3.2×10-13

30.0×10-13<Test conditions>(1) Corrosion resistance test against organic acid
・Test fluid: Organic acid with 2% concentration
・Temperature: 50°C
・Test time: 100 hours(2) Corrosion resistance test against sulfur
・Test fluid: Gasoline with 300 ppm of sulfur
・Temperature: 80°C
・Test time: 300 hours
Test pieces were immersed in the test fluid at theset temperature and left for the predetermined time.
The rate of weight change was calculated from theweight difference before and after the test.

4. Self-lubricating sintered bearings
Sintered oil retaining bearings for automobiles andoffice equipment may be used in environments withhigh temperatures or enclosed by a cover, etc. In thiscase, lubricant oil retained in the bearings may
evaporate and contaminate the surroundings. The
developed self-lubricating sintered bearings use
special graphite as the solid lubricant and areapplicable in high temperature environments where oil
cannot be used.
Regular graphite is inferior to other kinds of metal
powder in fluidity; however, increasing the additive
amount causes inferior compactibility preventing
formation of complex shapes and low green compact
strength.
Therefore, the metal powder mixture was changed
to adopt high-fluidity special graphite in an increased

Table 1 Result of corrosion resistance test
Rate of weight change %Developed material Cupronickel material(1) Corrosion resistance against organic acid

0.14

0.15

(2) Corrosion resistance against sulfur

0.10

0.40

-76-

*14E_*09 14/08/01 11:17 ページ 4

The Introduction of Sintered New Products for Automobile

The friction coefficient of the self lubricating
bearings were approximately half that of the bronze
series without oil content. The specific wear amount
was also about 1/3, showing an excellent anti-wear
property. This means they can be used in areas wherean anti-wear property is required without oil content.

amount. This achieved the development of self
lubricating sintered bearings with excellent sliding andanti-wear properties even when lubricant oil is not
used, while maintaining high productivity (Fig. 7).

5. Summarya) Sliding material for automobile parts

In this article, the newly developed highdensity/high-strength sintered machine parts forautomobiles and sintered bearings for special
environments were introduced.
We will promote marketing of the developed
products introduced in this article, along with
continuing development of higher functionality.

b) Bearings for office equipment

Fig. 7 Self-lubricating BEARPHITE

4.1 Structure and features of self lubricating
materials
Special graphite with improved fluidity was used as
the self lubricating material. The amount of graphite
was also increased by 5 times the regular bronzematerial, which was done by optimizing the blend ratioof the material powder. Table 3 shows the chemicalcomposition of the developed material and regularbronze series material.

[High-density/high-strength sintered material]・Achieved true density ratio of 95% to iron andYoung’s modulus equivalent to SUJ2
・Approximately 1.5 times radial crushing strengthand 2 times fatigue limit compared with theconventional high-density sintered materials[Corrosion resistance sintered bearings]
・Equivalent or better corrosion resistance and friction
coefficient and around 1/10 of specific wear amountof the cupronickel sintered bearings

Table 3 Chemical component of self-lubricating BEARPHITE
Chemical composition %Material

Cu

Sn

C

Developed productRemainder

7〜11

4〜17

Bronze Series

Remainder

7〜11

1〜12

[Self lubricating sintered bearings]
・Approximately 1/2 the friction coefficient andapproximately 1/3 the specific wear amountcompared to the bronze series sintered bearings
without lubricant content

4.2 Friction/wear properties
Table 4 shows the friction coefficient and specificwear amount without oil content.

ferences
Table 4 Friction coefficient and specific wear rate1) Sokeizai Center, Monthly Bulletin “Sokeizai” May2013 Vol. 54, No. 5 Special Issue, Sokeizai
Yearbook 2012

Developed material Cupronickel material
Friction coefficientSpecific wear rate m3/(N・m)

0.2〜0.3

0.45

0.8×10-14

2.9×10-14<Test conditions>(1) Friction test
・Load: 98N・Rotation speed: 1,000min-1
・Test piece size: ID 6 × OD 12 × Width 6 mm
・Shaft material: SUJ2
・Test time: 90min.
(2) Anti-wear test
・Load: 15N・Rotation speed: 450min-1
・Test piece size: ID 6 × OD 12 × Width 6 mm
・Opposite material: SUJ2
・Test time: 1 hour

Photo of authors-77-

Tomonori YAMASHITA

Tomokazu SONOZAKI

Engineering Dept.
NTN Powder Metal Corporation

Composite Material Engineering Dept.
Composite Material Products Division*15_*15 14/08/01 11:31 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Paper ]Rolling Contact Fatigue Life of Highly-carbonitrided JIS-SUJ2

Daisuke SATO*
Chikara OHKI*
Carbonitriding for rolling bearings has been known to prolong rolling
contact fatigue lives under both clean and debris contaminated lubrication
conditions. However, a relationship between surface nitrogenconcentrations and RCF lives has not yet been clarified. To this end, we
prepared carbonitrided rolling bearings with various concentrations ofsurface nitrogen. The kind of steel was JIS-SUJ2 (52100 and 100Cr6 equivalent). We conducted
three types of RCF tests; (1) with Rockwell indentations on the inner ring raceways, (2) with oil
containing hard steel particles, and (3) with clean oil. Test (1) emulates to evaluate RCF lives underdebris contaminated conditions.Test (1) exhibited the following results; the average RCF life of 0.4
mass % surface nitrogen concentration was about 2 and 3 times longer than that of conventional
carbonitrided and nitrogen free, respectively. Both tests (2) and (3) also demonstrated that an RCF
life with 0.4 mass % surface nitrogen concentration was longer than that of the others. Hence, it has
been proven conclusively that carbonitriding with high surface nitrogen concentrations brings about
excellent bearing performances.

1. Introduction

2. Surface initiated flaking life test

The impact of the carbonitriding process to
subsurface initiated flaking life 1) of rolling bearings
has been known for a long time by the study ofKurabe, et al. It was also recently confirmed, by
Dommarco, et al. about the longevity effect by an2), 3). For rollingincreased amount of retained austenite
bearings used in the transmission, hub and reducer of
automobiles, service life in contaminated lubrication(hereafter, surface initiated flaking life) is critical. It is
known that the carbonitriding process with anincreased amount of retained austenite is effective
under this condition, as well 4), 5). On the other hand, alife test of SUJ2 material with strict control of surface
nitrogen concentration has only been conducted by
Ohki, for the surface initiated flaking life 6) and no otherdata has been reported, yet.
In this report, we are introducing various evaluation
results of SUJ2 material with strict control of surface
nitrogen concentration. We have evaluated surface
initiated flaking life, lubrication life of clean oil andindentation resistance, which are critical properties forthe rolling bearings.

2.1 Bearing life test under artificially-applied
indentation
A bearing life test with artificial indentations(hereafter, artificial indentation life test) is conductedby forming artificial indentations on the raceway of theinner ring of the bearings, in advance, with clean
lubrication and is an effective method to appropriately
determine the surface initiated flaking life 6).

2.1.1 Types and materials of bearings
We used deep-groove rolling bearings (bearing
number: 6206, d = 30 mm, D = 62 mm, B = 16 mm, Z= 9) as test pieces. Table 1 shows the chemicalcomposition of JIS SUJ2, the material of the raceway
ring. The carbonitriding process of the inner ring was
conducted while monitoring partial pressure of CO and
CO2, as well as undecomposed NH3 with an infraredanalyzer.
Since the outer ring with no indentations will always
have longer life than the inner ring, the carbonitriding
process was not applied, but only ordinary quenching
and tempering. The heating temperature was set to
850°C for both and the tempering was conducted at a

*Advanced Technology R&D Center-78-

*15_*15 14/08/01 11:31 ページ 2

Rolling Contact Fatigue Life of Highly-carbonitrided JIS-SUJ2

The size of the indentations was about 200 μm in
diameter and 15 μm in depth.
Fig. 3 shows the schematic drawing of the life
tester. This structure allows the application of a radialload evenly onto two test bearings. Table 3 shows thetest conditions. An oscillation monitor was used for
detecting flaking. The oscillation acceleration of
normal operation was 1.5 m/s2. Therefore, it wasdetermined flaking had occured when it reached 6
m/s2, four times the normal operation, and the test was
terminated.

temperature of 180°C for two hours.
After randomly selecting one test piece each fromamong the pieces of ordinary quenching andcarbonitriding process, the carbon and nitrogenconcentration distribution at the center of the bottom of
the raceway of the inner ring was measured using theline analysis of the electron probe micro analyzer(EPMA). Fig. 1 shows the results. Based on thenitrogen concentration distribution, material was
removed from the surface of the inner ring raceway toachieve the target surface nitrogen concentrationlevel. The radial clearance of the test bearings was set
to be identical by adjusting the outer ring groove
diameter.
Table 2 shows the measurement results forretained austenite by X-ray diffraction. On the table,“0.4 mass % N” and “0.1 mass % N” mean the test
pieces with the surface nitrogen concentration of 0.4
mass % and 0.1 mass %, respectively, after
machining the test pieces.
The order of the test pieces with a high retained
austenite amount was the same as the order of test
pieces with a high concentration of nitrogen.
Fig. 2 shows the cross-sectional hardness
distribution of the test pieces after heat treatment.
There was no significant variation in the crosssectional hardness within the range of the heat
treatment conditions in this test. From the above, we
concluded that the key differences among these test
pieces are surface nitrogen concentration and theretained austenite amount.

Table 2 Retained austenite measured by X-ray diffraction
Material

Retained austenite amount at 0.05 mm deep
from scraped surface (%)

0.4mass%N

29.6

0.1mass%N

20.1

Ordinary tempered material

9.0

Hardness HV

800

750

700
0.4mass%N

650

0.1mass%N
Ordinary tempered material

600

0

0.10

0.20

0.30

0.40

Distance from the surface mm

2.1.2 Test method

0.50

Fig. 2 Hardness distribution profiles

The subject of the test is the artificial indentations
formed at the center of the bottom of the inner ring
raceway. These indentations were formed by pressinga Rockwell hardness testing penetrator made of conicaldiamond with a force of 196N. A total of 30 indentations
were equally spaced 12° apart per inner ring.

Load spring
Support bearing 6312
Test bearing 6206

Test bearing 6206

PulleyTable 1 Chemical compositions of JIS-SUJ2 used (mass%)
C

Si

Mn

P

S

Ni

Cr

Mo

Cu

Al Ti (ppm) O (ppm)

1.9
1.7

Ordinary temperedmaterial

1.5
1.3
1.1
0.9
0.7

Carbon

0.5
0.3
0.1
-0.1
0

Nitrogen0.1 0.2 0.3 0.4

0.5 0.6 0.7 0.8

Distance from the surface mm

Concentration of carbon and nitrogen %Concentration of carbon and nitrogen %0.99 0.24 0.40 0.012 0.006 0.06 1.36 0.03 0.10 0.011 0.24 0.04

1.9
1.7

Fig. 3 Schematic drawing of rolling contact fatigue life
tester for ball bearing

Carbonitriding
material

1.5
1.3
1.1

Table 3 Test conditions of artificial indentation life test

0.9
0.7
0.5
0.3
0.1
-0.1

0

0.1 0.2 0.3 0.4

CarbonRadial load kN

NitrogenRadial internal space µm

20Contact pressure Pmax GPa

3.1

Rotational speed of inner ring min-1

3,000

0.5 0.6 0.7 0.8

Distance from the surface mm

Lubricant oil

Fig. 1 Carbon and Nitrogen concentration distribution
profiles after heat treatment

Calculated life h

-79-

6.86

Turbine oil (VG56)
127.6

*15_*15 14/08/01 11:31 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

2.2.1 Types and materials of bearings

The life in this test was obtained assuming that it
followed the two-parameter Weibull distributionfunction of the Lundberg-Palmgren theory 7). Fig. 4
shows the Weibull distribution of the artificial
indentation life test and Table 4 shows the test results.
In terms of L50 life, the SUJ2 ordinary quenchingmaterial showed 69.0 h, the 0.1 mass% N material
showed 121.3 h and the 0.4 mass% N material
showed 223.7 h. This indicates that the higher thesurface nitrogen concentration, the longer the life, bothfor L10 and L50.
This test corresponds to the follow-up test of theprevious report 6). Similar to the previous report, it wasverified that the surface initiated flaking life could be
extended with high nitrogen concentration.

We used tapered roller bearings (bearing number:
30206, d = 30 mm, D = 62 mm, T = 17.25 mm, Z = 17)
as test pieces. The surface nitrogen concentration of
the raceway and rolling elements of the 0.4 mass % N
test piece after the machining process was all 0.4
mass %. Off-the-shelf product was used as theSCr420 carbonitriding test piece.

Cumulative probability of damage %2.1.3 Test results

2.2.2 Test method
Fig. 5 shows the schematic drawing of the life
tester. This structure allows the application of a radial
and axial load evenly onto two test bearings. Table 5shows the test conditions. The test was conducted byapplying a predetermined load on the test bearings,
injecting approximately 30 ml of contaminated oil into
the housing and rotating the inner ring of the bearings
at 2000 rpm. The detection method of flaking was thesame as the artificial indentation life test in 2.1.2.

99
80

Load spring

50

Pulley20
10Support bearing
Test bearing
30206

0.4mass%N
Ordinary tempered material

5

Ordinary tempered material6)
0.1mass%N6)

1
101

10

103

2

Life hFig. 5 Schematic drawing of rolling contact fatigue life
tester for tapered roller bearing

Fig. 4 Weibull distributions of artificial indentation life test

Table 4 Results of artificial indentation life test
Number oftest pieces

L10 life

L50 life

Weibull slope

0.4mass%N

8

119.9

223.7

3.5

0.1mass%N6)

12

81.7

121.3

4.8

Load kN

Ordinary temperedmaterial

12

40.6

69.0

3.6

Contact pressure Pmax GPa

2.5

Rotational speed of inner ring min-1

2,000

Material

Table 5 The conditions of life test under contaminatedlubrication
Fr

17.64

Fa

1.5

Lubricant

2.2 Contaminated lubrication life test
The artificial indentation test mentioned above
verified that the surface initiated flaking life of SUJ2
material with high surface nitrogen concentration could
be extended. In order to verify if the same trend canbe obtained even in actual contaminated lubricant, we
conducted a contaminated lubrication life test, which isthe conventional test method of the surface intiatedflaking life. In addition, we also compared performancewith the conventional SCr420 carbonitriding material.

Amount of contamination

1.0 g/LType of contamination

Gas-atomized powder
equivalent to KHA30
Size: 108 - 180 µm
Hardness: similar to HV800

Calculated life h-80-

Turbine oil (VG56)
Oil supply, approx. 30 ml

168.8

*15_*15 14/08/01 11:31 ページ 4

Rolling Contact Fatigue Life of Highly-carbonitrided JIS-SUJ2

3.1 Types and materials of bearings

Fig. 6 shows the Weibull distribution of the test
results. Table 6 shows the test results. In terms of L50
life, the conventional SCr420 carbonitriding material
showed 37.5h, SUJ2 ordinary quenching material
showed 70.3 h and the 0.4 mass % N material showed
149.1 h. Similar to the artificial indentation life test, it
was revealed that life could be extended by applying
the carbonitriding process to the inner/outer rings androlling elements. It is noted that the bearings with asurface nitrogen concentration of 0.4 mass % had acomparable service life in contaminated lubricant tothat of the calculated life of the SUJ2 ordinary
tempered material in clean lubrication.

The same test bearings as in Section 2.2 were
used.

Cumulative probability of damage %2.2.3 Test results

3.2 Test method
The tester in Fig. 5 was used. Table 7 shows thetest conditions of the life test in clean lubrication. The
test was conducted applying a predetermined load on
the test bearings, supplying oil in the housing androtating the inner ring at 2000 rpm.

3.3 Test results
The result of 7 test pieces of 0.4 mass% N showedno flaking after 7300 hours. Fig. 7 shows the Weibull
distribution. Arrows indicate suspension.
The relation between the suspended time and Ln
8)life is indicated in Eq. (1) .

99

｛

80

Ln = T50

10
0.4mass%N
Ordinary tempered material
SCr420 carbonitriding material

1
101

102

103

Life h｝Setting the target life to L10 and confidence level to90%, substituting termination time T of 7300h, thenumber of test pieces N of 7, n=10, Weibull slope e of
9/8 9), which is a typical value for rolling bearings, tothe formula (1), we obtained L10 life is 2600 h or morewith 90% reliability. This is 15 times longer than thecalculated life under clean lubrication, and isconsidered to be sufficient.

20

5

e
l（
n 1−0.01C）…………………（1）N（
ln 1−0.01n）

Fig. 6 Weibull distributions of life test under contaminatedlubrication
Table 7 The conditions of life test under clean lubrication
Table 6 Results of life test under contaminated lubrication
Material
0.4mass%N
Ordinary temperedmaterial
SCr420 carbonitriding
material

Number oftest pieces
6

L10 life
107.7

L50 life
149.1

Load

Weibull slopekN

Contact pressure

5.8

Fr

17.64

Fa

1.5

Rotational speed of inner ring
6
6

44.0
17.7

70.3
37.5

4.0

2.5

Pmax GPa
min-1

2,000
Turbine oil (VG56)

Lubricant

168.8

Calculated life h

2.5

Cumulative probability of damage %Note) The ordinary tempered material comes from a separate lot

3. Life test in clean lubrication
From the test result mentioned above, it was
revealed that the indensurface initiated flaking life canbe extended by increasing the surface nitrogenconcentration. However, it is necessary to verify
longevity in clean lubrication as the bearings aremostly used in the clean lubrication environment
without contamination.

99
80
50
20
10
5
1
101

0.4mass%N
Ordinary tempered material

10

2

10

3

Life h104

105

Fig. 7 Weibull distributions of life test under clean oil-81-

*15_*15 14/08/01 11:31 ページ 5

NTN TECHNICAL REVIEW No.81（2013）

In the figure, for example, “240°C-0.4 mass % N”
means that the tempering temperature is 240°C andthe surface nitrogen concentration after finishing work
is 0.4 mass %. The depth of indentation for 180°C 0.4
mass % N is less than SCr435 carbonitriding material
This is presumed to indicate an increase in the staticload-carrying capacity.
The depth of indentation tends to be less when thetempering temperature is high and more when thesurface nitrogen concentration is high. In general, theindentation becomes deeper when nitrided; however,
we have discovered that if tempered at 240°C after
nitriding, the depth of indentation becomes less thanthe ordinary SUJ2 quenching material.

4. Indentation formation test
All the results so far revealed that the surface
initiated flaking life and clean lubrication life can beextended by increasing the surface nitrogenconcentration. Next, we conducted an indentation
formulation test, which is to evaluate the depth ofindentation formed by an impression of a ceramic ballonto the test piece at a certain pressure in order toevaluate the static-load-carrying capacity.
The size of indentation in this test is very small witha depth typically 1 µm or less, unlike the artificial
impressions mentioned earlier, formed with stressequivalent to yield stress.

4.1 Shape of test piece and test method

5. Consideration

Fig. 8 shows the shape of the test piece. The
indentation is formed by pressing a silicon nitride
ceramic ball of 9.525 mm in diameter onto the mirrorfinished test piece surface. The depth of theindentation was measured with a three-dimensionalsurface shape measuring device.

5.1 Life test
This study revealed that the surface intiated flaking
life and the clean lubrication life can be extended byapplying a high nitrogen concentration on SUJ2. In
this section, we consider the main factors that
contributed to the extension of surface initiated flaking
life of the 0.4 mass % N materials. The increased
amount of retained austenite and solid solution
strengthening of nitrogen can be considered as thefactors. In order to confirm the impact of these factors
to the life of the bearings, we conducted an artificial
indentation life test with the high nitrogenconcentration 240°C 0.4 mass % N material with theequivalent amount of retained austenite as theordinary SUJ2 quenching material, and the ordinary
quenching 240°C tempering material with 1/2 or lessof retained austenite as the ordinary SUJ2 quenchingmaterial.
L10 life of the 240°C 0.4 mass % N material was
47.4 hours and L50 life was 87.7 hours, which areequivalent or slightly longer than the ordinary SUJ2
quenching material. L10 life of the ordinary quenching
240°C tempering material was 17.5 hours and L50 life
was 46.0 hours, which are shorter than the ordinary
SUJ2 quenching material. From these results, it isconsidered that the retained austenite amount, as wellas the solid solution strengthening of nitrogen affectthe surface initiated flaking life.

4.2 Test results
Fig. 9 shows the relation between the maximum
contact pressure and the depth of the indentation ofeach test piece. In this test, tempering temperature
was also changed in addition to the surface nitrogenconcentration.

4

φ40

Fig. 8 Shape of test piece with indentation

Depth of indentation μm

1.2
1.0
0.8
0.6

180˚C-0.4mass%N
210˚C-0.4mass%N
240˚C-0.4mass%N
180˚C-0.1mass%N

Ordinary tempered material
SCr435 carbonitriding
material

0.4

5.2 Prediction of the depth of indentation
0.2
0
3.0

3.5

4.0

4.5

5.0

Pressure GPa

5.5

Bearings may withstand instantaneous excessive
load if the material is resistant to indentation, evenwhen a heavy load is applied. Therefore, we exploreddesign guidelines for bearing material that withstandheavy load by understanding the relation between thematerial properties and the depth of indentation.
A silicon nitride ceramic ball of 9.525 mm in

6.0

Fig. 9 Relation between maximum contact pressure anddepth of indentation-82-

*15_*15 14/08/01 11:31 ページ 6

Rolling Contact Fatigue Life of Highly-carbonitrided JIS-SUJ2

Table 8 Depth and material properties of various test pieces
Depth ofindentation(actual measurement)

Retained
austenite(vol%)

SCr435 carbonitriding material

1.14

29.6

27.2

0.1

1.15

Ordinary tempered material

0.52

9.0

14.1

0.0

0.45

180˚C-0.1mass%N

0.63

20.1

12.7

0.1

0.67

180˚C-0.4mass%N

0.85

29.6

12.3

0.4

0.88

210˚C-0.4mass%N

0.79

21.9

12.3

0.4

0.71

240˚C-0.4mass%N

0.31

7.6

12.3

0.4

0.39

The coefficient of determination doubly adjusted forthe degrees of freedom R**2 in the estimation equation
is 0.92, and it was clarified that the retained austenite
amount and the average crystal grain diameter
contribute to the depth of indentation. The more theretained austenite amount and the larger the average
crystal grain size, the larger the indentation.
Therefore, material with refined crystal grains and areduced retained austenite amount is considered to besuitable for the bearings with excessive load.

diameter was pressed to the test pieces made of
SUJ2 ordinary quenching material, SUJ2
carbonitriding material and SCr435 carbonitriding
material with the load of Pmax=5 GPa and the depthof the indentations were measured. Table 8 shows thedepth of the indentations and properties of various
materials. Based on these results, we conducted
multiple regression analysis with the depth ofindentation as the dependent variable and the retained
austenite amount, average crystal grain diameter andsurface nitrogen concentration as the independent
variables. Since the surface nitrogen concentrationshows very little influence to the depth of indentation,
it was excluded from the independent variables for
multiple regression analysis. Equation (2) is theestimation equation for the depth of indentation
calculated from the multiple regression analysis andFig. 10 shows the relation between the actually
measured value and the estimated value of the depthof indentation.

6. Summary
In this paper, we introduced results of various life
tests with SUJ2 of high surface nitrogen concentration.
1) We conducted life tests with artificial indentations
on the inner ring raceway of rolling bearings 6206
for evaluation of surface initiated flaking life. In
terms of L50 life, the SUJ2 with high nitrogenconcentration showed approximately 3.2 times
longer life than the SUJ2 ordinary quenchingmaterial and approximately 1.8 times longer life
than the SUJ2 conventional nitriding material.
2) We conducted a life test of the tapered roller
bearings 30206 with high nitrogen concentration at
the inner/outer rings and rolling elements, incontaminated (with hard foreign objects) lubrication.
In terms of L50 life, the SUJ2 with high nitrogenconcentration showed approx. 2.1 times longer life
than the SUJ2 ordinary quenching material andapprox. 1.4 times longer life than the SCr435
carbonitriding material
3) As a result of the clean lubrication test of the SUJ2high nitrogen concentration material in the tapered
roller bearings 30206, none of the test pieces
experienced flaking over 7300 h, which, according
to the statistical calculation, indicates more than 15
times the calculated life with 90% reliability.
4) Regarding impact loads on the rolling bearings, weevaluated the likelihood of indentation by thecontact between the rolling elements and rotating

Depth of indentation (µm)
= -0.005 + 0.022 × retained austenite amount (vol%)+ 0.018 × average crystal grain diameter (µm)……(2)Measured depth of indentation μm

1.2

SCr435 carbonitriding
material

R**2=0.92
1.0

180℃-0.4mass%N
0.8

0.6

210℃-0.4mass%N
180℃-0.1mass%N

Ordinary temperedmaterial

0.4
240℃-0.4mass%N
0.2
0.2

Average crystal Surface nitrogen Depth ofgrain diameter
concentrationindentation(µm)(mass%)(estimate)Material

0.4

0.6

0.8

1.0

1.2

Estimated depth of indentation μm

Fig. 10 Measured and estimated values of depth ofindentation-83-*15_*15 14/08/01 11:31 ページ 7

NTN TECHNICAL REVIEW No.81（2013）rings. The result showed that the SUJ2 material
with high nitrogen concentration is less likely to beindented compared to the SCr435 carbonitriding
material, which is the current long-life material.
Increasing tempering temperature of the SUJ2
material with high nitrogen concentration further
improves this effect.

References
1) A. P. Voskamp：MICROSTRUCTURAL CHANGESDURING ROLLING CONTACT FATIGUE ,CHAPTER
3 (1996) 15.

2) Hyojiro Kurabe, Toru Araki: Tetsu to Hagane, vol.
53 (1967) 1305.
3) R. C. Dommarco, K. J. Kozaczek, P. C. Bstias, G. T.
Hahn, C. A. Rubin: Wear, 257 (2004) 1081.

4) Kikuo Maeda: Industrial Heating, vol. 38 (2001) 2.
5) C.Ohki : SAE Technical Paper Series (2004) 2004-010634.

6) Chikara Ohki: Tetsu to Hagane, vol. 95 (2009) 695.
7) Shigeo Shimizu: Kikaikei no tameno shinraisei
sekkei nyumon, Surikogakusha, Tokyo, (2006) 34.
8) Reliability Engineering Association of Japan:
Reliability Handbook (1997) 209.
9) G. Lundberg, A. Palmgren：IVA Handlinger,(1952)
210.

Photo of authors

Daisuke SATO

Chikara OHKI

Advanced Technology
R&D Center

Advanced Technology
R&D Center-84-

*16E_*15 14/08/01 11:34 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Paper ]Effect of Naturally Derived Antioxidants
as Surface Treatment
Yosuke TAGUCHI*
Eriko ITAHASHI*
Hidenobu MIKAMI*
Surface treatment is one of the techniques of improving thetribology characteristic and corrosion resistance. For example,
dry processings such as DLC and wet processings such as
plating and black oxide have been used to machine parts.
However, since some wet processings contain an environmental
pollutant in waste fluid, an environment-friendly surface treatment
is desired. In this paper, we introduce improvement of friction characteristic and bearing endurance
by a surface treatment which used "natural derived antioxidants" as processing agent.

1. Introduction

2. Natural antioxidant

Recently, many fields require efforts to reduce
environmental impact. Surface treatments, such as
plating and chemical treatment, have been
traditionally used for superior tribology characteristics
and corrosion resistance. However, these types of
surface treatments may discharge regulated
substances in the waste water.
Therefore, research and focus is on environmentallyfriendly surface treatments using plant-derived, natural
antioxidants. As an example tannic acid, which iscontained in persimmon, etc. is known to produce athin film coating on steel surfaces 1), 2), 3). However,tribological characteristics of this type of surface
treatment have hardly been evaluated.
In this report, the impact of surface treatment isevaluated, using chlorogenic acid, the plant-derived
natural antioxidant contained in coffee beans, etc. in
relation to various bearing performance characteristics
such as friction and high temperature. In addition, thesurface of the chlorogenic acid film was analyzed andthe chemical structure and composition areinvestigated.

Natural antioxidant is a collective term for thecompounds (polyphenol) containing multiple phenolichydroxyl (hydroxyl bound with aromatic rings such asa benzene ring and a naphthalene ring) in themolecule, and botanical component such as
carotenoids with C40H56, in the basic structure.
Natural antioxidants are abundant in green and yellow
vegetables. Fig. 1 shows the classification of natural
antioxidants. Curcumine, quercetin, isoflavone,chlorogenic acid, carotene and astaxanthin are wellknown to be effective as antioxidants when introduced
into human bodies 4).
In this article, chlorogenic acid was chosen as atreatment agent which is both soluble in water andexhibits relatively fast reaction with steel. Chlorogenic
acid is a compound contained in coffee beans (5-10%)
and has a structure made by dehydration
condensation of quinic acid and coffee acid, a
structural formula of which is shown in Table 1.

*Advanced Technology R&D Center-85-

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NTN TECHNICAL REVIEW No.81（2013）

Polyphenol
Naturaanti-oxidant

Curcuminoid

Curcumine

Flavonoid

Quercetin
Isoflavone

Chlorogenic acid
Carotenoid

Agitation impeller

Carotene
Astaxanthin

Fig. 1 Classification of naturally derived antioxidants

Chlorogenic acidsolution 0.5 wt %SUJ2 plate

Table 1 Structural formula of surface treatment
Name

Fig. 2 Schematic view of chlorogenic acid treatment

Structural formula
OH
O

HO

O

OH

O

Treatment
time

Chlorogenic acid

0 (no treatment)

1h

4h

8h

HO
HO

OH

Quinic acidAppearance

O

HO

OH

HO
HO
OH

Treatment
time
OH

Coffee acid
HO

O

OH

Appearance

3. Treatment method
10mm

Fig. 2 shows the chlorogenic acid treatment
schematic. The SUJ2 plate and steel strip used fordeep groove ball bearing rings and cages werecleaned with alkali and acid, immersed in the 0.5 wt %solution of chlorogenic acid, and stirred under room
temperature for a predetermined time.
Fig. 3 shows the appearance of the SUJ2 plateafter the chlorogenic acid treatment.
The color of the SUJ2 plate becomes darker the
longer the treatment time, indicating that a film on thesteel surface is produced.

Fig. 3 Chlorogenic acid treated SUJ2 plate

4. Evaluation of various performance
properties
Friction characteristics, high temperature durabilityof greased bearings and corrosion resistance of the
SUJ2 material treated with chlorogenic acid were
evaluated.

4.1 Friction property
Friction coefficient of SUJ2 plates with no treatment
and with 1, 4 and 8 hours of chlorogenic acidtreatment and 11/16 inch SUJ2 steel ball with notreatment was measured using a ball-on-disk typelinear wear friction tester. Fig. 4 shows the schematic
overview of the test and Table 2 shows the test
conditions.
Li soap/ester grease was used as the lubricant.
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Effect of Naturally Derived Antioxidants as Surface Treatment

4.2 High temperature durability of greased bearings

Fig. 5 shows the results of the friction coefficientmeasurement. At the beginning of the test the friction
coefficient of the chlorogenic acid treated plate was
the same as that of the plate without treatment;
however, it started decreasing soon thereafter with afinal value 0.02 lower than the plate with no treatment.
It is also noted that the chlorogenic acid treated plateshowed a lower friction coefficient than the plate withno treatment regardless of the treatment time.
The results showed no difference in wear amountbetween the untreated sample and that with thechlorogenic acid treatment.

Using a tester complying with ASTM D 3336,
greased deep groove ball bearings incorporating steel
cages treated with chlorogenic acid were evaluated fordurability.
Steel cages with no treatment, 1-hour treatment and4-hour treatment with chlorogenic acid wereincorporated into the 6204 ZZ bearings (ID 20 mm x
OD 47 mm x Width 17 mm) with a dimension
stabilization process. Fig. 6 shows the overview of thetest equipment and Table 3 shows the test conditions.
The test was conducted three times for each treatedcage and average lives of n=3 were evaluated.
Fig. 7 shows an appearance of steel cage treatedwith chlorogenic acid and Table 4 shows scanningelectron microscope (SEM) photos and arithmetic
average roughness (Ra) of the chlorogenic acidtreated cage surface. As shown, the roughness of thecage surface increased as the treatment timeincreased.
Fig. 8 shows the results of the high temperature
durability test. The bearing using the 1-hour cagetreatment did not show any increase in life relative tothe standard part. The bearing using the cage treatedfor 4 hours showed more than twice the life of the
bearing with no cage treatment.

Reciprocating motion

Steel ballGrease

SUJ2 plate

Fig. 4 Principal part schematic view of reciprocation test

Table 2 Reciprocation test condition
Plate test piece

HeaterBearings
under test

SUJ2 (φ24, t=8)0.02 〜 0.05

Plate surface roughness Ra
Opposite material

Coil spring

11/16 inch SUJ2 steel ballGreaseSupportbearings

Li soap/ester grease

Frequency

10 Hz

Temperature

40 ˚CAmplitude

1.2 mmLoad

15 N

Maximum contact pressure

0.8 GPa

Test time

30 minPulleyto Motor

Thermocouple

Fig. 6 High temperature bearing durability tester
0.30
No treatment

Table 3 High temperature bearing durability test condition

1h

Friction coefficient

0.25

4h
8h

0.20
0.15
0.10
0

5

10

15

20

25

30

Test time min

Fig. 5 Chronological change of friction coefficient-87-

Bearings

Deep groove ball bearings 6204ZZ
ID 20 mm x OD 47 mm x Width 17 mm

Sealed grease

Li soap/ester grease

Amount of sealed grease

0.71 gTemperature of bearings

150 ˚C

Rotational speed

10,000 min-1

Radial load

67N

Axial load

67N

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NTN TECHNICAL REVIEW No.81（2013）

4.3 Corrosion resistanceCorrosion resistance was tested under high humidity
conditions compliant to JIS K 2246. Fig. 9 shows theschematic overview of the test and Table 5 shows thetest conditions. SPCC material of 80 × 60 × 1.2 mm ofsize, treated with chlorogenic acid for 4 hours was usedas the test sample. For comparison, test pieces with notreatment and blackening process on the steel surface
were also used. These test pieces were hung under thetest stand in a dry condition without oil or grease suchas rust-preventive oil, and exposed to the temperatureof 49°C and relative humidity of 95% or more. The test
pieces were examined for rust every 24 hours.
Table 6 shows the results of the corrosion resistancetest under high humidity conditions. The test piece withno treatment produced rust within the first 24 hours.
After one month, little rust was observed on the test
piece treated with chlorogenic acid; however, after 2
months rust appeared. The test piece with blackening
treatment did not produce rust even after 3 months. It
was found that, although inferior to the blackening
process , the chlorogenic acid film is effective against
corrosion.

Fig. 7 Cage after 4 hours of chlorogenic acid treatment

Table 4 SEM images of steel retainer surfaces
Treatment Ra, µm
time, h0
(no treatment)

SEM images

0.15

100µm

Test piece stand

0.15

1

Test piece

Air

Pure water

Fig. 9 Schematic view of high humidity corrosion test

100µm

Table 5 High humidity corrosion test condition

0.26

4

Item

Test condition(1) Rotation speed of test piece stand

0.33mm-1(2) Temperature of test piece position

49±1 ˚C(3) Relative humidity

95％% or moreThree times of
the tank capacity/h(4) Air flow
100µmLife against seizure h

600(6) Water

500

Pure water

Table 6 High humidity corrosion test result

400

Surface treatment

300

Test time

No treatment

Blackening

Chlorogenic acidtreatment

24h

×

○

○

200
100
0

Adjusted to conform tothe conditions (2) and (3)(5) Temperature of water tank

0

1

2

3

4

5

6

Treatment time hFig. 8 High temperature bearing durability test result

1 month

×○○

2 months×○△

3 months×○○：Not rusted △：Partially rusted-88-

×
×：Fully rusted

*16E_*15 14/08/01 11:34 ページ 5

Effect of Naturally Derived Antioxidants as Surface Treatment

C-O

5. Surface analysis
8000

The chlorogenic acid film exhibited friction reduction
effect, extended life against grease seizure andcorrosion resistance. In order to find the reasons forthese effects, surface analysis of the film on thetreated SUJ2 plate was performed.

Strength cps

7000

5.1 XPS analysis

0 min

6000

5 min

5000

10 min4000

15 min

3000
2000
1000

The chlorogenic acid treated surface was examined
for the chemical composition and structure using X-ray
photoelectron spectroscopy (XPS).

0290

5.1.1 Qualitative analysis (survey analysis)

288

286

284

1.2
1.0
0.8
0.6
0.4
0.2
0
1100 1000 900

Strength cps

Strength cps

O KLL

C1s

O1s

280

278

Fig. 11 C1s spectrum of cholorogenic acid film

Fe2p3

Fe2O3

C-O

9000

1.6
1.4

282

Binding Energy eV

Fig. 10 shows the chemical composition of thesurface of the SUJ2 plate treated for 4 hours. C, O
and Fe were detected on the film surface.

X104
2.0
1.8

C-C
C-Fe

8000

0 min

7000

5 min

6000

10 min5000

15 min

4000
3000
2000
1000
0
544

800

700

600

500

400

Binding Energy eV

300

200

100

540

0

536

532

Binding Energy eV

528

524

Fig. 12 O1s spectrum of cholorogenic acid film

Fig. 10 XPS spectrum of film surface

5.1.2 Depth and composition analysis

Fe
C-Fe

Fe2O3

50000
0 min
40000

Strength cps

Fig. 11 to 13 show the narrow spectra of C, O and
Fe on the surface of the 4-hour treated SUJ2 plate.
The sputter-etching times indicted in the figures are 0,
5, 10 and 15 min. and the etching speed is equivalent
to the SiO2 standard test piece's 30 nm/min.
The C1s spectrum showed a change of peak
positions; C-O and C-C bonding were most prevalenton the upper most surface which transitioned to C-Fe
bonding as the depth increased (Fig. 11). The O1s
and Fe2p3 spectra revealed that Fe on the upper
most surface was Fe2O3, and that while the depth
increased, C-Fe and Fe also increased (Fig. 12, 13).

30000

5 min
10 min15 min

20000
10000
0
730

725

720

715

710

Binding Energy eV

705

Fig. 13 Fe2p3 spectrum of cholorogenic acid film

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NTN TECHNICAL REVIEW No.81（2013）

5.2 TOF-SIMS analysis

Fig. 14 shows the composition ratio on the surfaceof the SUJ2 plate treated for 4 hours. C was mostprevalent on the upper most surface while little Fe wasobserved. As the depth increased, the ratio of C
decreased while Fe increased.
Fig. 15 shows C concentration is dependent on thetreatment times. C was only detected on the upper
most surface of the plates with no treatment and 1
hour of treatment; however, C was detected even after
about 25-30 minutes of sputtering on the plates
treated for 4 or more hours. After about 25-30
minutes of sputtering with this XPS equipment, theetching depth is about 0.7-0.9 µm.
The XPS analysis revealed that the chlorogenic acidfilm consists of C, O and Fe and is about 0.7-0.9 µm
thick. The uppermost layer consists of a C-richcompound including C-O and C-C, as well as a small
amount of Fe2O3. As the depth increases, C and O
decrease and Fe increases. Fe2O3 found on thesurface layer decreases and compound with C-Fe
bonding increases.

As the XPS analysis results show that the surface
layer of the chlorogenic acid film is an organic
substance including C-O and C-C bonding, TOF-SIMS(time of flight secondary ion mass spectrometer)
analysis was conducted.
Fig. 16 shows the chlorogenic acid agent mass
spectrum.
From the main peak (m/z=353), chlorogenic acidions (C16H17O9-) were detected.
Fig. 17 shows the mass spectrum of the SUJ2 plate
treated for four hours with chlorogenic acid.
Chlorogenic acid ions (C16H17O9-) equivalent tom/z=353 were not detected from Fig. 17 (a). However,as shown in Fig. 17 (b) and (c), fragment ions similar
to quinic acid (C7H9O4Fe-) and coffee acid
(C9H7O5Fe-) were detected.
Therefore, the upper surface of chlorogenic acid film
is partly formed by quinic acid and coffee acid which
are decomposed from chlorogenic acid.
While finding the mechanism of reduced friction
remains, it is assumed that the low friction is the resultof the “conformability effect” of the formation of C-richsoft film including quinic acid and coffee acid which
are decomposed after the chlorogenic acid treatment.

Atom concentration %

100
80
C1s

60

O1s

40

Fe2p3

20
0

0

10

20

30

40

Sputter-etching time min

Fig. 14 Compositional ratio of a chlorogenic acid film

Atom concentration %100

Notreatment

80

1h
C1s
4h
O1s
8h
Fe2p3

60
40
20
0

0

10

20

30

40

Sputter-etching time min

Fig. 15 Effect of chlorogenic acid treatment time-90-

*16E_*15 14/08/01 11:34 ページ 7

Effect of Naturally Derived Antioxidants as Surface Treatment

Strength cps

X104
2.0
1.5
1.0
0.5
320

340

360

380

400

Mass m/z

420

440

460

480

Fig. 16 Mass spectrum of a chlorogenic acid agent

Strength cps

X104

(a)

2.0
1.5
1.0
0.5
320

340

360

380

400

420

440

460

480

Mass m/z

Strength cps

X102

(b)

0.8
0.6

Quinic acid

0.4
0.2
212.0

Quinic acid
212.5

213.0

Mass m/z

213.5

214.0

(c)

1.0
0.8

Coffee acid

0.6

Strength cos

Strength cps

X102

0.4
0.2
250.0

250.5

251.0

Mass m/z

251.5

252.0

Fig. 17 Mass spectra of chlorogenic acid treated SUJ2 plate

References

6. Summary

1) Yoshiaki SHIBATA, Takayuki SHIROMOTO, Makoto
YUASA, Isao SEKINE, Toshinobu IMAHAMA, Toshiharu
WAKE: Corrosion Inhibition of Mild Steel by Gallic Acid in
Neutral Solution, Journal of the Surface Finishing Societyof Japan, 44, 4 (1993) 347
2) Masayoshi KATO, Takakazu KURODA, Daisuke
YAMAMOTO: Kinzoku Hyomen Kogyo Zensho 13, Metal
corrosion resistance technology (Maki Shoten, 1969)
3) Japan Society of Corrosion Engineering: Handbook ofCorrosion Technologies (Boshoku Gijutsu Benran)(Nikkan Kogyo Shimbun, Ltd., 1986)
4) Hoyoku NISHINO, “Development of Free Radical Theory
and Food Products for Ageing Prevention”, CMC
Technical Library, pp. 14-21 (1999)

It was found that by applying chlorogenic acid as
the surface treatment, the frictional properties,
greased bearing life and corrosion resistance are
improved.
After four hours, the thickness of the chlorogenic
acid film on the treated SUJ2 plate will become
uniform. It was also found that the upper most surfaceof the film is organic and the compound of C-Fe
increases as the depth increases. In addition, thepresence of decomposed substances of chlorogenic
acid such as quinic acid and coffee acid was found on
the film surface.
In addition to the characteristics introduced in thisarticle, the chlorogenic acid treatment is expected to
be used as a high functionality treatment that saves
energy and extends life. The waste water after
treatment only includes chlorogenic acid, thus isenvironmentally friendly.
The mechanism of various characteristics thatchlorogenic acid brings will be investigated in order toreduce environmental impact by applying this surface
treatment for various applications.

Photo of authors (titles are at the time of development)-91-

Yosuke TAGUCHI

Eriko ITAHASHI

Hidenobu MIKAMI

Advanced Technology
R&D Center

Advanced Technology
R&D Center

Advanced Technology
R&D Center*17E_*15 14/08/01 11:46 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Paper ]Grease Leakage Prevention
for Double-Row Angular Ball Bearings

Mitsuo KAWAMURA*
Norihide SATO*
Tomoya SAKAGUCHI*
Mariko IZUMI**
Grease leakage from sealed bearings filled with grease makes bearing
lives shortened and the peripheral parts polluted. The grease leakage
generally occurs between relatively rotating inner seal grooves and seal
rips. Hence, the authors had developed shape modified ribbon cages
which suppress the grease transfer to the seal grooves and improved
grease leakage prevention performance of ball bearings.
By adapting the approach of grease leakage prevention for ribbon cages to snap cages, greaseleakage prevention performance of double-row angular ball bearings was examined. Theseexperimental results revealed that the bearings with the modified snap cages have excellent greaseleakage prevention performance. In addition, structure analyses of the cage were conducted and themodified snap cage was confirmed to be strong enough in practical use.

Applications with outer ring rotation which frequently
start and stop are the most susceptible to greaseleakage 4), 5). A typical application example is the useof double-row angular contact ball bearings in
automobile electric auxiliary machines. When greaseleaking from the bearings contacts the belt of theauxiliary machine, the belt can slip and cannot
sufficiently drive the electric auxiliary machine.
Because of this grease leakage must be controlled by
the bearings.
The above double-row ball bearing is equipped with
two snap cages inserted from either side of the
bearing due to the structural limitation of the bearing
with two rows of raceway grooves.
In this report, we attempted to reduce the greaseleakage of double-row ball bearings by applying thegrease leakage prevention technology that
established with ribbon cages to snap cages.

1. Introduction
Grease leakage from sealed and greased bearings
may cause shorter overall bearing life andcontaminate the surroundings. Countermeasures suchas increased force of seal lip on the inner ring seal
groove 1) and modification of seal lip shape 2) havebeen adopted; however, these changes cause anincrease in rotational torque and higher costs.
Reducing the initial grease amount is also effective
for reducing grease leakage; however, because thisapproach shortens grease lubrication life, it is normal
practice to seal as much grease as possible within theextent that the leakage does not occur 3) .
Since the grease of bearings leaks from between
the inner ring seal groove and seal lip, we devised aribbon cage with a special shape capable ofcontrolling the grease around the inner ring seal
groove. We confirmed its effectiveness in reducing
grease leakage 4), 5).

**Advanced Technology R&D Center**CAE Dept., Automotive Business HQ-92-

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Grease Leakage Prevention for Double-Row Angular Ball Bearings

III) When the grease Go on the outer surface of theinner ring increases, some grease attaches to thefacing inner surface of the cage.
However, this location is the center of the pocket,
most expanded on the axial direction, resulting in
the grease attachment Gtr shown in Fig. 1 c).
IV) When the grease Gtr increases, a part of thegrease is pushed up to the inner ring seal groove
and is attached as shown by Gg in Fig. 1 d). When
this continues, the grease is attached to the entire
circumference of the inner ring seal groove.
Fig. 2 shows the standard ribbon cage and the
ribbon cage with grease leakage prevention capability
developed previously. It also shows how the grease isattached when the outer rings of the bearings
equipped with these cages are rotated. Fig. 2 a)shows a standard ribbon cage with which a lot ofgrease attaches to the inner seal groove when theouter ring rotates. Fig. 2 b) shows a ribbon cage withgrease leakage prevention capability where two areasof the inner surface of the pocket are recessed. With
this design, grease does not attach to the inner ring
seal groove, even when the outer ring rotates. We
also verified that the grease leakage from the bearings
is significantly reduced when operated with the seal.

2. Grease leakage prevention
technology for ribbon cages 4), 5)
Since grease leakage of bearings occurs between
the inner seal groove and seal lip, it is necessary thatgrease attach on the inner ring seal groove. Withouter ring rotation (the condition most susceptible togrease leakage), the movement of grease from therolling elements to the inner ring seal groove has been
observed to be as follows (Fig. 1):
I) The grease attached to the rolling elements iswiped off by the inner side edge of the cage pocket
and attaches to the edge of the pocket as Gwr in
Fig. 1 a).
II) When the amount of grease Gwr at the edge of thepocket increases, some grease attaches to theouter surface of the inner ring as shown by Go in
Fig. 1 b).

a) Wiping stepb) Grease transfer step to outer surface of inner ring
Grease on inner ring seal groovea) Conventional ribbon cagec) Grease transfer step to pocket center area

No grease on inner ring seal grooved) Grease shift step to inner seal grooveb) Cage of recessed pocket at diagonal positions

Fig. 1 Grease migration from ball surface to inner ring seal
grooveFig. 2 Shape of cages and location of grease of ballbearings after outer ring rotation

-93-

d

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NTN TECHNICAL REVIEW No.81（2013）

bearings. For calculating the grease interface position,
the volume of fluid method, which calculates interfaceof mixed-phase fluids, was used. Since the bearings
have periodicity and symmetry, the domain for thisunsteady analysis was focused on one pocket only.
The fluids were assumed to be incompressible andmove with laminar flow. For simplicity, the temperatureand density of fluid was set constant and a small
space was necessary between the balls and raceway
due to constraints in the calculation method.
Fig. 5 shows an example of the analysis on the
migration of attached grease as it relates to therotation of the bearings. The red and yellow colors
both indicate the grease inside the bearings; yellow
indicating the grease attached on the seal surface.
The analysis result in Fig. 5 also indicates a similar
trend as Fig. 4 d). This analysis also verifies that thegrease attached to the balls was wiped off by the edgeof the cage pocket, accumulated on the inner surfaceof the cage, and after increasing over time, migrated
to the edge on the closed side of the cage.
From the results of the above observation andanalysis, the migration path of the grease to the inner

3. Grease behavior with snap cages
With double-row angular contact ball bearings, theclosed side of the snap cage is always oriented to
face of the outboard side of the bearings, as shown in
Fig. 3. Therefore, it is assumed that the greaseleakage can be prevented if the grease migration fromthe closed side of the snap cage to the seal groove iscontrolled.
Counter-opening side of pocket

Outer ring

Cage insert
Seal

Cage insert
Steel Ball

Cage

Inner ring

Fig. 3 Cross section view of double-row angular bearings

We examined the grease behavior of the closed
side of snap cages in 6203 size deep groove ballbearings. For enhanced visibility of grease location,
the grease was colored red and the cage was coloredwhite. The grease thickener used was urea-basedwith worked penetration of 280. After sealing 0.5 g ofgrease on the raceway, the outer ring was rotatedclockwise at an ultra-low speed under 7.8 N of axialpressure.
Fig. 4 shows the grease behavior at the outer ringrotational angles of 0, 180, 360, 720, 1080 and 1800
degrees. In Fig. 4 b) and c), grease is observed
between the cage and the raceway ring, indicatingthat the grease attached to the rolling element was
wiped off by the cage at the very beginning of therotation. In Fig. 4 d), grease is observed at the edgeof the closed side of the snap cage. In Fig. 4 e),grease is attached at the boundary between the outersurface of the inner ring and the seal groove and in
Fig. 4 f), this grease is increased, starting to attach tothe seal groove.
The attachment of grease to the seal groove
mentioned above is assumed to be triggered by theattachment of grease on the edge of the closed side of
the cage shown in Fig. 4 d). Therefore, the grease
behavior was examined using computational fluid
dynamics (CFD) analysis in order to verify the
behavior of grease attachment on the edge of theclosed side of the cage.
In the CFD analysis, fluid with rheological propertiesof grease and air were placed in the space inside theOuter ring rotation

Load direction on inner ringa) 0˚b) 180˚
Grease on cage outside
Grease on the cage edge
Grease on cage borec) 360˚d) 720˚

Grease on inner seal groovee) 1080˚f) 1800˚
Fig. 4 Grease behavior in a deep groove ball bearing on
the counter-opening side of the cage pocket-94-

*17E_*15 14/08/01 11:46 ページ 4

Grease Leakage Prevention for Double-Row Angular Ball Bearings

III) The above processes I) and II) repeat along withthe rotation, and the grease attaches to the entire
circumference of the inner surface of the cage andouter surface of the inner ring as the grease Gis and
Go in Fig. 6 c).
IV) Furthermore, by the increase of the grease Gws
and the grease pushed out to the axial direction
between the raceway and balls, the grease Gis and
Go are pushed out and attach to the seal groove as
the grease Gg shown in Fig. 6 d). Since the cage
and the inner ring rotate relatively, the grease Gg
spreads out on the entire circumference.
When the axial load was applied on the opposite
direction, there was no significant change in thegrease behavior; attachment of grease was observedon the seal groove of the inner ring.

seal groove on closed side of the bearing cage wasestimated as in Fig. 6.
I) Part of the grease attached to the rolling elements iswiped off by the inner surface of the cage andaccumulates as the grease Gws shown in Fig. 6 a).
II) The grease Gws wiped off by the inner surface ofthe cage increases and attaches to the outersurface of the inner ring (grease Go ) as shown in
Fig. 6 b).
Grease on the cage edge

4. Snap cage with grease leakage
prevention capability

BallFig. 5 Result of numerical simulation of grease behavior
in deep groove ball bearings

Compared to the standard snap cage shown in Fig.
7 a), we devised a cage with a modified shape(hereafter, “shape modified cage”) as shown in Fig. 7b), based on the estimation of the grease migration
path presented in the previous Section and the resultof Section 2. We provided recessed areas, similar toOpen sidea) Wiping stepAb) Grease transfer step to outer surface of inner ringa) Conventional cagec) Growing step of grease Gis and Go

B

ConcaveCd) Grease move step to inner seal groove
Fig. 6 Migration path of grease from ball surface to innerseal groove in ball bearings with snap cagesb) Shape modified cage
Fig. 7 Snap cage shape-95-

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NTN TECHNICAL REVIEW No.81（2013）

Table 2 Conditions of grease leakage test

the ribbon cage, on the inner surface of the pocket ofthe shape modified cage. The intention is to reduce
the wipe-off effect on the inner surface of the cage.
Since the grease tends to accumulate on part A of
Fig. 7, it was removed as shown in part B of Fig. 7 b).
Rigidity is not affected by the amount of material
removed. In order to ensure the strength of the cage,an annular reinforcement was added at the bottom of
the pocket where the cross sectional area is thesmallest, as shown in part C of Fig. 7 b). The strengthof the cage is reviewed in Section 7.
The cylinder-shape protrusions in the axial direction
between the pockets are provided for a manufacturing
reason, when the bearings are assembled.

Double-row angular bearing
with contact rubber seals

Bearing type
Bearing size, mm

φ35×φ50×20

Outer ring rotation speed，min-1

13,500

Axial load，NTemperature

1,470

Test time，h

2

R.T.

6. Evaluation result of grease leakage
prevention property and consideration
6.1 Grease behavioral test result
Fig. 8 shows how the grease behaved after therotation of the outer ring in the grease behavioral test.
A lot of grease was observed on the closed side of thestandard cage in Fig. 8 a). This is considered to be
due to the large amount of the grease Gws in Fig. 6,
which was wiped off by the inner surface of the cage.
Fig. 8 b) shows the test result when the shapemodified cage was incorporated. Only a little amountof grease was attached on the inner surface of thecage, and no grease was found on the closed side of
the cage pocket, outer surface of the inner ring, orseal groove.

5. Evaluation method for greaseleakage prevention capability
In order to evaluate performance of the designed
cage, we verified the effect on the grease behavior
and the prevention capability against grease leakage
from the bearings.

5.1 Grease behavioral test
In order to confirm the change of the grease
behavior due to the shape modification of the cage,
we operated the double-row angular ball bearings
under the conditions listed in Table 1 and observed
how the grease attached on the inner ring seal
groove.

Grease on edge of cage bore

Table 1 Conditions for confirming shape modified cage effect
Bearing type
Bearing size, mm

Double-row angular bearinga) Double-row angular bearing with conventional cage

φ35×φ50×20

Outer ring rotation speed，min-1

1,800

Rotating time，sAxial load，N

39.2

10

It was necessary to observe the difference in thegrease behavior from the counter-opening side of thecage pocket; therefore, grease was injected from theopening side of the pocket before the operation.
To realize this, the raceway was cut in the middle ofaxial direction and used as the single-row angular ballbearings.

b) Double-row angular bearing with shape modified cage

Fig. 8 Location of grease after outer ring rotation

6.2 Grease leakage test result
Fig. 9 shows the relationship between the grease
ratio and the reduction of mass from the greaseleakage test. The solid (filled-in) plots in the figure
indicate the bearings which grease leakage wasvisually observed after the test. The grease ratio is theratio of grease volume to the space capacity where
grease can remain static even when the balls and the

5.2 Grease leakage test
In order to quantitatively evaluate the greaseleakage prevention capability in the bearings, weverified the leakage amount depending on various
greased amounts. Table 2 shows the test conditions.
The grease used in this test is the same as that usedin Section 3.
-96-

*17E_*15 14/08/01 11:46 ページ 6

Grease Leakage Prevention for Double-Row Angular Ball Bearings

speed variation of balls”) and a relatively large force isapplied onto the cage 6), 7) by balls.
Therefore, we examined the strength of the cages
against centrifugal force and orbital speed variation ofballs.

cage rotate in the bearings. When this ratio exceeds
100%, grease is constantly stirred and likely to leak.
Fig. 9 reveals that the standard cage allows greaseleakage at 65% or more, while the shape modified
cage retained grease up until 90% of the grease ratio.
The minimum mass reduction amount that thegrease leakage could be visually detected was about
0.07 g. The limit of grease ratio that does not allow
grease leakage of 0.07 g or more is 65% for the
standard cage and 90% for the shape modified cage.
This example shows that the limit of the greased
amount could be increased by 25% by employing theshape modified cage. *From the results mentioned above, it is evident thatthe shape modified cage improves the grease leakage
prevention property, as well as increases the amountof grease in the bearings without allowing leakage.

Fig. 10 shows distribution of maximum principal
stress due to centrifugal force on the standard andshape modified cages. While the maximum principal
stress on the standard cage was shown at the inner
Low stressHigh stressAmount of visible grease leakage

0.10

Weightloss of bearing g

7.1 Strength of cage against centrifugal force

Inner side

Maximum stress

0.08
0.06
0.04

Conventional cage
Shape modified cage

0.02

Filled plot : grease leakage isvisually inspected

0.00
50

60

70

80

90

Ratio of filled grease to static volume %Maximum stress100

Inner side

Fig. 9 Result of grease leakage testa) Conventional cage

* The space for grease to remain static is smaller for the shapemodified cage compared to the standard cage because of thelarger cross-sectional area at the bottom of the pocket.
However, the difference of this static space volume is only
2%, the negligible level in Fig. 9.

Low stressHigh stressMaximum stress

7. Examination of cage strength
Inner side

It was assumed that the magnitude of stress anddisplacement during operation would change due tothe difference of shape between the shape modified
cage and the standard cage. Therefore, we examined
the strength of the cage by completing a static
structure analysis using a finite element method.
The protruded edges of the pocket of the snap cageare pushed outward when the bearings rotate at highspeed. This creates concern that the protruded edge andthe outer ring will contact, as well as a resulting increaseof contact load between the ball and the pocket.
Additionally, when the bearings become misaligned,
the orbital speed of the balls varies (hereafter, “orbital

Maximum stress

Inner sideb) Shape modified cage

Fig. 10 Maximum principal stress of cages due to centrifugal force-97-

*17E_*15 14/08/01 11:46 ページ 7

NTN TECHNICAL REVIEW No.81（2013）

surface side of the bottom of the pocket, the maximum
principal stress on the shape modified cage wasshown at the outer surface side. This change of the
maximum principal stress point location is due to thechange of the deformability of the entire cage by theaddition of two recessed areas on the inner surface
and removal of some material between pockets.
Fig. 11 shows the summary of main principal stresson these cages. The maximum principal stress on the
shape modified cage decreased to one half comparedto the standard cage. This is due to the increase of thecross-sectional area at the bottom of the pocket, in
addition to the change of balance in deformability,
mentioned above.
Fig. 11 also shows the maximum deformation,
which reveals that the maximum deformation of the
shape modified cage decreased to one half comparedto the standard cage.
Since deformation of the protruded edges on theouter surface, where the maximum deformation
occurs, is affected by the strength of the bottom of thepocket, we consider that the reinforcement of part C ofFig. 7 greatly contributed to this change.

Speed of faster ball

1.0

Speed of slower ballContact loads by ballsa) Loads acting on cage due to orbital speed variation of balls

Displacement constraint

Constraint of circumferential direction

Conventional cage
Shape modified cage

1.2

Ratio

variation of balls. The cut section on the left is theconstrained surface and the load from the ball isapplied to the right side of the central pocket. As
shown in Fig. 13, this force due to the orbital speed
variation of the balls created the maximum principal
stress at the bottom of the pocket for both cages.
However, the maximum principal stress points areeach different; the standard cage showed theb) Boundary condition of cage stress analysis

Fig. 12 Boundary condition of cage stress analysis due
to orbital speed variation of balls

0.8
0.6

Low stress

0.4

High stress

0.2
0.0

StressDeformation

Fig. 11 Ratio of maximum stress and deformation of
cages due to centrifugal force

Cut section
Maximum stress

7.2 Strength of cage against orbital speed
variation of balls

Inner sidea) Conventional cage

Fig. 12 a) shows the operational behavior of the balland the cage when the orbital speed of the balls
varies. The balls with higher orbital speed and lower
orbital speed make contact with the pocket surface,respectively, creating a strain state between bothpockets. In order to analyze the stress on the cage,
we applied periodically symmetrical boundary
conditions on the cross sections and forced
displacement equivalent to the contact load from theballs to the end pocket surfaces in the half
circumference of the cage.
Fig. 13 shows distribution of the maximum principal
stress generated in the cage due to the orbital speed

Maximum stress

Cut section
Inner sideb) Shape modified cage

Fig. 13 Maximum principal stress of cagedue to orbital speed variation of balls-98-*17E_*15 14/08/01 11:46 ページ 8

Grease Leakage Prevention for Double-Row Angular Ball Bearings

maximum stress at the inner surface while the shaped
modified cage showed it at the outer surface. This isdue to the change in the balance of deformability.
When operating bearings with large moment load,
cages may break at the bottom of the pockets and the
maximum principal stress was found at the same area
as these breaking points.
Fig. 14 shows the ratio of the maximum principal
stress. It reveals that the maximum principal stress ofthe shape modified cage decreased by 28%
compared to the standard cage.
From these analyses, the stress generated in the
shape modified cage is lower than that in the standard
cage, even when centrifugal force under high rotationor force from the balls due to orbital speed variation isapplied. The shape modified cage possesses
sufficient unit strength in practical applications.

with this technique. The following are the results:
1) The observation of grease behavior revealed thatthe shape modified cage can reduce the greaseamount attached to the inner surface of the cage
and as a result, reduce the grease migration into
the inner ring seal groove of double-row ballbearings.
2) The double-row ball bearings with the shapemodified cages did not produce grease leakage
even when the grease ratio over static spacevolume of bearings was 90%. On the other hand,the standard cage produced leakage with 65% of
the grease ratio over static space volume.
3) The maximum principal stress generated on the
shape modified cage due to centrifugal force
decreased to one half that of the standard cage.
The deformation of the protruded edge from thestress also decreased to half. The maximum
principal stress of the shape modified cage
produced by the orbital speed variation of the balls
was 28% lower than that of the standard cage.
Because of this, we believe that the shape modified
cages have sufficient strength for practicalapplications.

1.2

Stress ratio

1.0
0.8
0.6
0.4
0.2
0.0

References
Conventional
cage1) Nozaki, Okasaka, Kubota, Akabe: Trends in Automotive
Instrument and Auxiliary Bearing Technology, NTN
TECHNICAL REVIEW No. 65, (1996), 65
2) NTN Ball Bearings/Shield and Seal Types, Catalog: CAT. No.
3015-III/J, (2000), 3
3) Tomobe, Uematsu: Grease Amount in Shield Bearings,
Journal of Japan Society of Lubrication Engineers, 19, 4, 261,
19744) Sato, Sakaguchi: Improvement of Grease Leakage Prevention
for Ball Bearings Due to Geometrical Change of Ribbon
Cages, NTN TECHNICAL REVIEW, 78, 98, 2010.
5) Sato, Sakaguchi: Reduction of Grease Leakage Prevention
for Ball Bearings Due to Geometrical Change of Cages,
Tribologist, 57, 4, 281, 2012.
6) Kakuta: The Forces Acting on the Retainer of Ball Bearings
under Radial or Thrust Load, Transactions of the Japan
Society of Mechanical Engineers (Part 3), 32, 1161, 1966.
7) Kakuta: The Forces Acting on the Retainers of Ball Bearings
under Moment Load or Misalignment, Transactions of theJapan Society of Mechanical Engineers (Part 3), 32, 1176,
1966.

Shape modified
cage

Fig. 14 Ratio of maximum principal stress of cage due toorbital speed variation of balls

8. Summary
We applied the cage design method developed forribbon cages to the snap cages for double-row ballbearings and evaluated grease leakage prevention
properties. Additionally, we analyzed the stress on thecage due to centrifugal force from high speed rotationand due to orbital speed variation of the balls causedby misaligned bearings using the finite element
analysis techniques. Both the strength of the shapemodified cage and the standard cage were analyzed
Photo of authors

Mitsuo KAWAMURA

Norihide SATO

Tomoya SAKAGUCHIIMariko IZUMIAdvanced Technology
R&D Center

Advanced Technology
R&D Center

Advanced Technology
R&D Center

CAE Dept.
Automotive Business HQ-99-

*18E_*15 14/08/01 11:54 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ Technical Paper ]A Creep Mechanism of Rolling Bearings

Tsuyoshi NIWA*

The mechanism of creep phenomena, in which a bearing ring gradually
moves circumferentially to the stationary shaft or housing mating with thering, has not been clarified yet. The author found a “surface travelingwave creep” mechanism based on a numerical analysis and theoreticalconsideration. The creep is featured by movement in the same direction
as the bearing rotation. In addition to the creep mechanism, this paper
proposes a design requirement to prevent the creep.

experimentally, Soda, by pointing out an example of acanister with a rubber band around it, as shown in Fig.
2, explained that creep occurs “because a small
slippage between the rubber band and the tea
canister at a certain point below the load accumulates
moving forward as the canister rolls on.” At the
beginning, the rubber band deforms by the load of thetea canister and stretches out in the circumferential
direction. This deformation is symmetrical and theslippage is also identical between the right and left
sides, therefore creep, which is a movement to one
direction, should not occur. However, when the tea
canister rolls, the deformation caused by the previous
slippage still remains, causing asymmetry. Creep isthen produced as the slippage to one direction
increases. The rubber band creeps to the opposite
direction from the rolling direction of the canister.

1. Introduction
1.1 Creep in the opposite direction of rotation
Soda 1) describes two mechanisms regarding thecreep phenomenon of the bearing ring in the opposite
direction of rotation.
One of them is relatively well knownand is causedby clearance between the ring and the housing or
shaft, as shown in Fig. 1. Namely, when the rotationalload acts on the bearings, the ring rotates on the fitting
surface of the housing based on the change of theload direction. If the clearance between the fitting
surface and ring is c, the ring delays by πc per rotationof the bearing, causing creep in the opposite direction
from the rotation of the bearing.
The other one is caused by elastic deformation andslippage of the ring when there is no clearance. While
Imai 2) also proved presence of this mechanism

Clearance c

Rolling
element

Direction of load

Ring

Direction of rotational load

Rubber bandTeacanisterDirection of load
Traveling directionof canister
Rotational directionof canister

Direction of bearing rotation
Table

Creep direction of ring
Housing

Fig. 1 Creep by clearance

Creep directionof rubber band

Fig. 2 Creep by rubber band slip

*CAE Dept., Automotive Business HQ-100-

*18E_*15 14/08/01 11:54 ページ 2

A Creep Mechanism of Rolling Bearings

1.2 Creep in the same direction as rotation

2. Mechanism of traveling wave typecreep

On the other hand, actual bearings often creep inthe same direction as rotation. This phenomenon canbe explained by indicating the mechanism for the ringto creep in the same direction as the traveling
direction of the rolling elements.
Ten et. al 3) explains that “We consider that force torotate the outer ring on the circumference wasproduced from the strain that occurred every time therolling elements traversed, which pushed the outerring forward, little by little, generating creep.” Fig. 3
shows the schematic diagram.
This mechanism, explaining that creep comes fromthe accumulated strain moving forward on thecircumference, matches with the explanation of Soda
and can be considered to reflect the actual
phenomenon. However, while Sakajiri et al. 4) pointed
out some interesting remarks on the wavy deformationof the outer ring surface, the explanation of Ten et al.
does not give mention to thisfinding and therefore,
does not give a clear picture of the phenomenon.

Direction of load
Rotational direction of bearings
Creep direction of ring

2.1 Transfer of ultrasonic motor
We look at the operation of an ultrasonic motor. In a
disk-type ultrasonic motor, traveling waves are excitedby piezoelectric elements on the surface of an elastic
body. This produces an elliptic motion on the rotor in
the opposite direction of the traveling waves on thesurface of the elastic body 5).
Fig. 4 shows the transfer mechanism of theultrasonic motor. The rotor, pressed on the elastic
body, moves in the opposite direction from thetraveling waves by the elliptic motion of the surface.

Pressing force

Direction of traveling waves

Direction of rotor movement

Elliptic motion of surface

Rotor

Rotational directionof rolling element
Traveling directionof rolling element

Stator (elastic body)Fig. 4 Mechanism of ultrasonic motor

Stretching of ring
Shrinkage of ring

2.2 Self-transfer of ring due to the traveling waves

Rolling
element

According to Sakajiri et al. 4) the traveling waves inbearings are produced when the load from the rolling
elements acts on the ring surface and the fitting
surface immediately below that point is pushed out.
This causes the surface to become wavy up to
several microns depending on the load. When the
bearings rotate, the rolling elements also move, withwhich these surface waves become traveling waves.
In the ultrasonic motor application, the travelingwaves produced at the stator move the rotor. With thebearing ring creep, as shown in Fig. 5, travelingwaves are produced to the ring itself which cause theit to creep. In terms of the direction of creep, the ringacts on the housing in the opposite direction of the

Ring

Housing

Fig. 3 Creep by strain

1.3 Proposal from our study
In this study, we provided a hypothesis that thetraveling wave is produced on the ring surface andthis traveling wave moves the ring and produces
creep. Using an FE analysis, creep can be produced
relatively easily. However, there is currently noliterature pointing out that creep is a transfer
phenomenon by traveling waves. This may be causedby the fact that this phenomenon is unique and the
explanation of the behavior itself is difficult. Therefore,
we will start explaining this mechanism using familiar
devices. Then, we will also discuss conditions to
restrict creep from the principles of generating
traveling waves.

Load of rolling element

Traveling directionof rolling element

Rotational direction of bearings

Elliptic motion of surface

Creep direction of ring

Ring

Rolling
element

Housing

Fig. 5 Traveling-wave creep of bearing ring-101-

*18E_*15 14/08/01 11:54 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

3.2 Evaluation by Bussinesq equation

traveling direction of the rolling elements. However
since the housing is fixed, the ring is pushed backresulting in traveling in the same direction as therolling elements, that is, rotating in the same direction
as the rotation of the bearing.
The traveling waves produced on the ring surface
behave like peristaltic motion in the circumferential
and radial directions throughout the loaded area. The
rotational movement opposite to the traveling waves
at the fitting surface immediately below the rolling
elements is characteristic of this.
In this study, this transfer phenomenon of the ring by
this mechanism is called "traveling wave type creep."

We examined the behavior of the point in the bodyusing the Bussinesq’s displacement equation 6) whichassumes plane strain.
Fig. 7 shows the analysis model. Following themodel of Imai 2), we simplified the model assumingthat the load of the rolling elements qi is distributed
following the sinusoidal function. The displacement
produced to the point at the depth y in the middle ofthe loaded area affected by the i th load is describedby Equations (1) and (2). Fig. 8 shows the Lissajous
figure indicating the locus of the point in the bodywhich is displaced by the load of 5 traveling rolling
elements.

3. Conditions to restrict the travelingwave type creepLxi

Ten et al. 3) pointed out that increasing the numberof rolling elements for load distribution and using
thicker ring are effective for reducing the local strain.
However, conditions to fundamentally restrict creep
are not indicated. In the following sections, we will
discuss the design requirements necessary for
restricting the traveling waves.

qi

v

y
(2)

qi

-1 x ixi y− (1−2ν)tan
22yxi ＋ y

｛
q
x
v =
｛x y
2πG
ui =

As shown in Fig. 6, it is considered that a point
located at a relatively shallow area of a body moves ina loop when load from traveling rolling elements is
applied on the surface of the body. When a rolling
element passes immediately above (1), the point ispressed down and when the point is between tworolling elements (3), it moves up closer to the surface.
In between those positions (2) and (4), it moves to the
opposite sides of the rolling elements. This micro
behavior is repeated.

w

y

u

Fig. 7 Analytical model

3.1 Behavior of a point within a body

(1)

w

i

2πG
i

2

i

i ＋

（ ）｝ ……… (1)

2

2

（

＋ 2(1−ν)log

xi 2＋ y 2B）｝…… (2)

L : Length of load zone
qi : Load of rolling element ixi : Position of rolling element iu, v : Displacement
G : Shear modulusν : Poisson’s ratio
B : Integral Constantw : Pitch between rolling elementsinterval between
w：Pitchrolling elements
y：DepthAy① y/w =0.25(Magnification×1)(3)A② y/w =0.5

(Magnification×2)A③ y/w =0.63(Magnification×4)

Fig. 8 Locus of internal pointy
From Equations (1) and (2), if the ratio of the pitch w
and the depth y is constant, that is, if the conditions
are homologous, the shape and size of the loop in the
Lissajous figure become the same. For convenience,
B of Equation (1) was set to 1, however, it does notaffect the shape and size of the loop.

y

(4)

y
Fig. 6 Behavior of internal point under loading-102-

*18E_*15 14/08/01 11:54 ページ 4

A Creep Mechanism of Rolling Bearings

In this study, we considered that the movement of theelliptic motion in Fig. 6 and 8 is the source of the creep
movement by traveling waves and evaluated theamplitude A of Fig. 8, as the index of the elliptic motion.

4. Summary(1) We have clarified the phenomenon where thering creeps in the same direction as thebearing rotation and presented a mechanism
called "traveling wave type creep”.
(2) We have presented the mechanism to restrict
the traveling wave type creep under thecondition of the thickness of ring / pitchinterval of rolling elements > 0.6.

3.3 Condition to restrict the traveling wave
Fig. 9 shows the examined result of the behavior of
amplitude A in Fig. 8. While the amplitude A is
proportional to the sum of the rolling element load Σqi ,it converges under the conditions of the ratio of depth y/pitch interval w > 0.6.
The ring, housing and shaft are in contact,
therefore, the above discussion that assumes
continuous body does not generally apply. However, if
we consider the depth y as the thickness of the ring t,
under the condition of Equation 3 below,

In addition, as further complex contact behavior on
the fitting surface due to traveling waves is being
clarified 7), we are planning to study it in details.
This article has been edited based on the original
manuscript for the Proceedings of Tribology
Conference 8). We thank the Japanese Society of
Tribologists, who gave us the permission for
publishing it.

Thickness of ring t / pitch interval of rolling elements w > 0.6….….….….…. (3)the traveling waves, which is the main driving forceof the transfer phenomenon, converge. Therefore,
creep should be restricted and is depicted in Fig. 10.
From the above, we consider that the conditions for
producing traveling wave type creep are in the ratio of
the thickness of the ring and pitch interval rather thanthe load and size of the bearing components.

References
1) Soda: Bearings, Iwanami Shoten (1964), 196.
2) Imai: Lubrication, 4 (1959) 307.
3) Ten, Sakajiri, Takemura, Yukawa: NSK Technical
Journal No. 680 (2006) 13.
4) Sakajiri, Takemura: Proceeding of Japan Tribology
Conference (2003-5) 103.
5) Kanazawa, Tsukimoto, Maeno, Miyake: Tribologist,38, 2 (1993) 21.
6) For example, Nakahara: Zairyo Rikigaku - Gekan,
Yokendo (1966) 122.
7) Niwa: Proceedings of Japan Tribology Conference(2012-9) 141.
8) Niwa: Proceedings of Japan Tribology Conference(2012-5) 253.

Amplitude AΣqi=0.5Σqi=1.0Σqi=2.0

0.0

0.2

0.4

0.6

Depth y / pitch interval w

0.8

Fig. 6 Amplitude of locus

t/w＜0.6w

Photo of authorstt/w＞0.6t

Aw

Tsuyoshi NIWAA≒0

CAE Dept.
Automotive Business HQ

Fig. 10 Thickness of bearing ring and traveling-wave-103-

*19_22E_*18,19_受賞案件 14/08/11 16:29 ページ 1

NTN TECHNICAL REVIEW No. 81（2013）

Presentation of award-winning articles“The Japanese Society of Tribologists 2012” Technology Award

Prevention of Grease Leakage from Ball Bearings
Due to Cage Shape Improvement
Norihide SATO,

Mitsuo KAWAMURA,

Tomoya SAKAGUCHI

Table 1 Results of grease leakage test from sealed ball bearings

1. Introduction

Cage shape

Grease leakage in contact seal type, greased deep grooveball bearings occurs from between the inner seal groove andthe seal lip which rotate relative to each other. Therefore, we
considered that the grease leakage could be controlled bycontrolling the attachment of grease to these components. With
the development of a cage with a new geometry that could
control the attachment of grease to the inner ring seal groove,
we succeeded in improving the grease leakage prevention
properties 1), 2), 3).
This technology received the “2012 Japanese Society of
Tribologists Technology Award”, which is given to excellent newtechnologies in tribology. In this article, an overview of thistechnology is introduced.

Base cage
Press cage of diagonally recessed pocket innersurfaces (Fig.2 (a))
Press cage of narrow pocket width (Fig.2 (b))

0／15Weight loss of bearing g

0.06
0.04

Base cage
Developed cage

0.02

Filled plot: grease leakage isvisually inspected
50

60

70

80

90

100

Ratio of filled grease to static volume %%Fig. 3 Grease leakage characteristics of bearings with developed snap cagegrease amount using bearings that incorporated the snap cages
shown in Fig. 2 (c) demonstrated superior grease leakage
prevention property compared to the standard snap cages, as
shown in Fig. 3.

4. Summary
By using bearings with this technology, it is possible toextend bearing life by increasing grease amount while reducing
grease leakage. By extending life, bearings may be downsized,
leading to reduced rotational torque and lighter weight. While
this technology is currently adopted in automotive electrical andauxiliary equipment, we expect to expand the application tovarious machines and contribute to improved energy saving.

③④

Fig. 1 Path of grease transfer to inner seal groove

(a)

1／15

0.08

0

When the outer ring or inner ring of ball bearings rotate at low
speed, grease attached to the balls is wiped off by the cage andattaches to part ① of Fig. 1. Then the grease migrates to theouter surface of the inner ring ②, in the middle of pocket ③, andinner ring groove ④.
With this technology, we developed a cage with geometry
shown in Fig. 2, which can control the grease migration path
mentioned above.

②

6／15

0.10

2. Behavior of grease attachment to innerring seal groove①

Leakages／Total

References
1) Sato, Sakaguchi: Improvement of Grease Leakage Prevention for Ball
Bearings Due to Geometrical Change of Ribbon Cages, NTN TECHNICALREVIEW, 78 (2010) 91.
2) Sato, Sakaguchi: Reduction of Grease Leakage Prevention for Ball Bearings
Due to Geometrical Change of Cages, Tribologist, 57, 4 (2012), 281.
3) Kawamura, Sato, Sakaguchi: Improvement of Grease Leakage Prevention for
Double-Row Angular Ball Bearings, Proceedings of Tribology Conference,
Fukui, 2010-9, (2009) 283.

(b)

Photo of authors(C)

Fig. 2 Shape of developed cages

3. Evaluation result of grease leakage
prevention property
As shown in Table 1, the bearings incorporating the ribbon
cages of the geometry in Fig. 2 (a) and (b) significantly reduced
a frequency of the grease leakage compared to the standard
ribbon cages.
Furthermore, evaluation of the grease leakage with different-104-

Norihide SATO

Mitsuo KAWAMURA

Tomoya SAKAGUCHI

Advanced Technology
R&D Center

Advanced Technology
R&D Center

Advanced Technology
R&D Center*19_22E_*18,19_受賞案件 14/08/11 16:29 ページ 2

NTN TECHNICAL REVIEW No. 81（2013）

Presentation of award-winning articles
EMM2012 Mechatronics Award Grand PrixIntelligent In-wheel Motor System

Herve LENON*， Eiichi ASANO**， Tomomi GOTOU***，

Kaori TOMODA***run of “Q'mo”, which are the unique attributes of the in-wheel
motor drive, drawing the attention of many visitors.
At the international conference concurrently held, we gave
a technical presentation called “Optimization of energy
efficiency of EVs by applying in-wheel motor and by-wiresystem" to raise the presence of EV system products.
The meeting gives awards to excellent products every
year and NTN’s intelligent in-wheel motor system won thehighest honor of Grand Prix among the six nominatedcompanies.
This in-wheel motor system was highly praised for the
integration of innovative mechatronics technologies forcontrolling rotation of wheels, motor, reducer, inverter, brakeand related sensors, and was unanimously elected by thejudges.

1. Introduction
At the “European Mechatronics Meeting (EMM) 2012”
held in the Haute-Savoie Department of Eastern France from
June 6 to 7, 2012, NTN’s Intelligent In-Wheel Motor System
won the Grand Prix of Mechatronics Award.
EMM2012 is an international meeting in the field of
mechatronics organized by Thésame, a public industrial
support organization based in the Haute-Savoie Department.
The meeting attracted approximately 300 participants from
20 companies both inside and outside of France, featuring
exhibits and conferences of mechatronics technologies
related to environmental conservation such as naturalenergy, EV/HEV, etc.

2. Overview of presentation
3. Summary

NTN participated in the meeting together with ourEuropean subsidiary NTN-SNR, which is headquartered in
France, and appealed the technical alliance of Japan andFrance with the aim of acquiring EV-related business in theEuropean market.
At the booth, we presented a dynamic demonstration of“In-Wheel Motor System forElectrical Commuters” andvarious sensor technologies.
In the special exhibit area,
we exhibited a two-seater
electric commuter with the “InWheel Motor System”
installed on rear wheels and aconcept car “Q’mo” with the“Four-Wheel Independent
Drive/Steering System”
installed. We also
demonstrated “lateral move”
and “pivot turn” in the demo

We will promote early adoption of EV system products bypresenting and supporting the latest EV technologies of the
NTN group in the European market where EVs are becoming
popular.

***NTN-SNR ROULEMENTS Reserch & Innovation Mechatronics***Business Promotion Dept. EV Module Business HQ***Drive System Engineering Dept. EV Module Business HQ-105-

*19_22E_*18,19_受賞案件 14/08/11 16:29 ページ 3

NTN TECHNICAL REVIEW No. 81（2013）

Presentation of award-winning articles
Automechanika 2012 Innovation Award

In-wheel Motor System
Herve LENON*， Eiichi ASANO**

1. Introduction

3. Summary

The In-Wheel Motor System developed by NTN won the
Innovation Award in the parts division at “Automechanika
2012”, the world's largest exhibition for the automotive
aftermarket that was held in Frankfurt, Germany from
September 11 to 16, 2012.
Automechanika is the world’s largest exhibition for the
automotive aftermarket and held every other year.
In 2012, it attracted more than 4,500 companies from 74
countries and approximately 150,000 people participatedfrom all over the world.

In-wheel motors are gaining a lot of attention as thedriving system for the next generation electric vehicles withthe expectation for improved safety and vehicle stability as
well as reduction of greenhouse gas. We strive to expand
our activities globally for the early adoption of the next
generation EVs, including in the European market where
EVs are increasingly gaining ground.

2. Overview of presentation
NTN’s European subsidiary, NTN-SNR exhibited at
Automechanika 2012 with the theme "Innovation and R&D."
It presented a large range of aftermarket products such as
wheel bearings with sensors, engine timing belts, boots for
auxiliary equipment, auto tensioner, water pump kit, etc., aswell as the concept cars “Q’mo” and “Two-seater ElectricCommuter” equipped with an in-wheel motor system.
The most outstanding technologies at the show wereawarded for each division and NTN’s In-Wheel Motor
System won the Innovation Award in the part division.
With 13 European automobile related companies entering
the award category, the In-Wheel Motor System was highly
praised as being the world's lightest and most compact
design consisting of the hub bearing, motor, and reducer
stored within the wheel, and also for its outstanding
functionality as a vehicle stability control system thatprovides independent control of the left and right wheels
based on sensor information.
The show and the award presentation ceremony wereintroduced by TV, magazines, and the internet, which helped
to raise the presence of the NTN group in the automotive
aftermarket and EV market.

NTN-SNR representatives receiving the award
Ulrici, Vice President, Automotive Business HQ (2nd from left)
Gauthier, Sales Manager, Automotive Repair (Center)
Malavasi, Vice President, Automotive Repair/Logistics (2nd from right)**NTN-SNR ROULEMENTS Reserch & Innovation Mechatronics**Business Promotion Dept. EV Module Business HQ-106-

*19_22E_*18,19_受賞案件 14/08/11 16:29 ページ 4

NTN TECHNICAL REVIEW No. 81（2013）

Presentation of award-winning articles
The Japan Society for Precision Engineering, Spring Meeting, 2003, Best Presentation Award

Parallel Link High Speed Angle Control Equipment
Hiroshi ISOBE, Yukihiro NISHIO, Keisuke SONE, Hiroyuki YAMADA, Yoshio FUJIKAWA

1. Overview

Drivemechanism

NTN developed Parallel Link High Speed Angle Control
Equipment (hereafter, high speed angle control equipment) 1), 2), 3)
applying a special parallel link mechanism with rotational 2degrees-of-freedom, and deployed this equipment to a grease
dispensing mechanism. The key features of this equipment arehigh speed operation and wide range of motion.
The presentation 4) on this equipment given at the Spring
Meeting, 2003 of The Japan Society for Precision Engineering
received the “Best Presentation Award” for young researchers,
which is selected from the perspective of the originality of theresearch, presentation content, Q&A, etc. In this article, an
overview of this presentation is introduced.

Parallel linkmechanism

Tip-side link hub

Fig.1 Parallel link high speed angle control equipment

0°φ

2. About high speed angle control equipment

θ

Third link system

First link system

Fig. 1 shows the high speed angle control equipment
developed for grease dispensing. This equipment determines
the attitude of parallel link mechanism with 4 joints and 3 links
represented by the bend angleθand angle of traverseφin Fig.
2. The attitude is uniquely determined if two among three baseside arm rotational angles βn (n=1,2,3) are determined. The
relational expression is shown in Eq. 1.
cos(θ/2)sinβ1 - sin(θ/2)sin(φ+δ1)cosβ1 + sin(γ/ 2)=0
cos(θ/2)sinβ2 - sin(θ/2)sin(φ+δ2)cosβ2 + sin(γ/ 2)=0
cos(θ/2)sinβ3 - sin(θ/2)sin(φ+δ3)cosβ3 + sin(γ/ 2)=0…………………(1)We have improved the repeat positioning accuracy 2) up to±0.03mm or less, by controlling all 3 links by motors andallowing slight interference of 3 base-side arms for canceling thebacklash of the drive mechanism.

γ
Base-sidearm

Second link system
δ3

β1
β2

δ1=0°

δ2

Fig.2 Parallel link mechanism

High speed angle
control equipment
Stand
Work

3. Application to grease dispensing mechanism
An example of the grease application equipment
configuration using this equipment is shown in Fig. 3.
In this configuration, it moves the work in the XY plane, as
well as controls the attitude of the jet dispenser (a model todischarge grease at a distance by controlling high pressure air
with electromagnetic valve). With this configuration, grease canbe applied from various directions with different angles, to areas
previously not reachable from the top or vertical side of acolumn-like object. We have also achieved high speed
application of approx. 0.1 s/point.

Dispenser

XY stage

Fig.3 System configuration example of grease dispensing

Photo of author (representative)

References
1) Keisuke Sone, Hiroshi Isobe, Koji Yamada: High Angle Active Link
Equipment, NTN TECHNICAL REVIEW No. 71, 70-73, 2003
2) Hiroshi Isobe, Yukihiro Nishio: Parallel Link High Speed Angle Control
Equipment (PHACE), NTN TECHNICAL REVIEW, No. 80, 42-47, 2012
3) Hiroshi Isobe: Parallel Link High Speed Angle Control Equipment – Basic
Configuration and Application for Grease Dispensing, Machine and Tool,
February Issue, 83-87, 2013
4) Hiroshi Isobe, Yukihiro Nishio, Keisuke Sone, Hiroyuki Yamada, Yoshio
Fujikawa: Parallel Link High Speed Angle Control Equipment, Proceedingsof the Japan Society for Precision Engineering Spring Meeting in 2013,809-810, 2013

Hiroshi ISOBE
New Product Development
R&D Center-107-*23_26E_*20_26製品紹介 14/08/01 14:50 ページ 1

NTN TECHNICAL REVIEW No.81（2013）[ New Products Information ]High Load Capacity Cylindrical Roller Bearings
The world's highest load capacity and high speed rotating performancewith a rolling element separator. Longer operating life is achieved!

Rolling element separator

N typeNU typeFeatures

Structure[1] Has both high load capacity and highspeed rotating performance

Bearings with cages

Developed product• Rated life: 1.5 times of cylindrical rolling bearings
with cages• High speed rotating performance: 1.5 times of fullcomplement roller bearings

The number of rollers
is restricted for
securing strength ofcages

High load capacity
achieved bypreventing contact
of rollers using
the separator[2] Lighter and more compact• Optimal design of rolling elements to allow internalgear surfaces and rolling shafts to be used asouter ring track surfaces

Full complement
roller bearings

Large friction resistance dueto the contact of rollers incounter rotation

Applications
Load bearing capacity¡Wind turbine gearboxes and reducers ofindustrial machines

Bearings with cages

Low speed
shaftHigh speed
shaft

Main input shaft

Full complement roller
bearings

Permissible
rotation
speed

Resistancto smearing

Planetary
gear
Example configuration ofplanetary gear using
internal gear surfaces asouter ring track surfaces

Developed product

Planetary
section

Output,power
generatorside
Medium speed
shaft

Wind turbine
gearbox layout

Rated life-108-

*23_26E_*20_26製品紹介 14/08/01 14:50 ページ 2

New Products Information

Plastic Sliding Nut and Screw
with High Load Capacity
An industry first! A sliding nut and screw that integrates brass andplastic for superior durability and wear resistance!

Screw shaft

Combined nut

Features

Structure
Brass

By the hybrid structure of metal and resin,[1] Load resistance:
Twice compared to that of conventional
products[2] Durability:
Three times compared to that ofconventional products

PPS resin-based
sliding material[3] Capable of being used without lubrication:
Low noise, low heat generation

Test condition
Stroke: 180 mm
Wear after 2000 times: ≦0.1 mm

Applications
1,800¡Clean environment
Permissible load N(medical equipment/food processing equipment, etc.)

1,500

Developed product

1,200
900
600

Conventional product

300
0

100

200

Rotation rpm

-109-

300

*23_26E_*20_26製品紹介 14/08/01 14:50 ページ 3

NTN TECHNICAL REVIEW No.81（2013）

Multi Layer BEARPHITE
The world's first!
A double molding method to cast two dissimilar materials into a single object!!
Oil-impregnated sintered bearings with high strength and outstanding wear resistance.

Features

Structure

A double molding method is used to cast two
dissimilar materials into a single object to achieveboth high strength and outstanding wear resistance.

・High wear resistance material is used on the inner layer ofthe bearings and high strength structural material is usedon the outer layer:⇒achieves bearing performance and moldability
・Simultaneous molding and sintering of two dissimilar materials:⇒Strong bonding due to the interface diffusion during sintering[1] Radial crushing strength
Oil content: 17 vol% or more
[2] Variation in inner and outer diameter
dimensional precision:
50 µm or lessOuter layer
High-strength
structure

Applications

Inner
layer
内層軸受部
of bearings¡Bearings used for joints of construction
machines, etc.

High strength/high wear
resistance materia

part: applicable
Sintering of solid/liquid layers-110-*23_26E_*20_26製品紹介 14/08/01 14:50 ページ 4

New Products Information

Hybrid BEARPHITE HP
An industry first! Hybrid BEARPHITE that can be used in high PV ranges
10 times that of conventional products.

Features

Structure

By adopting ultrahigh molecular weight resin:High molecularresin materialHigh molecularresin material[1] Bearings are usable in high PV ranges
10 times or more than that of conventional products[2] Can be molded with the same dimensions as
rolling bearings:
Directly replaceable
Sintered metal[3] Low friction, high level of rotational
precision, compatible with flexible shafts:

Strong bondingby anchor effectHigh performance and can be used with aluminum or
resin shafts

Bearing size:φ8×φ16×5
Shaft material: Bearing steel
Test temperature: Room temperature

Applications¡Office equipment, photocopiers, bearings for LBP, etc.
Tray

Friction coefficient µ

×10 -8 mm/N/m
25

Development

Copying

Sintered metal

Exposure

15

Transportation-111-

Conventional material
Developed material

1050Paper feed tray

Applicable in high PV
range due to adoption ofultrahigh molecular resin

20

0

5

10

15

20

PV value MPa・m/min

25

30