Special Issue;
Automotive Environmental
Technologies
October 2007

ISSN 0915-0528 OSAKA,JAPANR

TECHNICAL REVIEW
No.75

Special Issue

corporation

Automotive Environmental
Technologies

NTN TECHNICAL REVIEW No.75
CONTENTS
Kenji OKADA

Preface

Contribution

Saving Energy with Motor Vehicle Transmission Technologies
Haruo HOUJOH TOKYO INSTITUTE OF TECHNOLOGY Precision and Intelligence Laboratory

1

2● Driveline
Technical Trends in Constant Velocity Universal Joints and the Development of Related Products
Shin TOMOGAMI

Fixed Constant Velocity Joint with a Super High Operating Angle of 54 Degrees (TUJ)
Manabu HOSHINO and Masashi FUNAHASHI

Measurement of the Internal Forces of Ball Fixed Constant Velocity Joints

10
16

Daiji OKAMOTO and Hirokazu OOBA

20

Technical Trends in Axle Bearings and the Development of Related Products

29

Kiyotake SHIBATA and Takayuki NORIMATSU

Development of High Resolution Sensor Element MPS40S and Dual Track Magnetic Encoder
for Rotational Speed and Position Measurement
Pascal DESBIOLLES and Achim FRIZ
Tube Forming Simulation in a Generation 4 Hub joint
Akitoshi IMOU and Takehiro TAKANODevelopment of an In-Wheel Motor Axle Unit
Minoru SUZUKI, Kayo SAKAI, Koichi OKADA and Yusuke MAKINO

Development of an Electromechanical Brake
Tatsuya YAMASAKI, Masaaki EGUCHI and Yusuke MAKINO

36
42
46
53

● Engine & Transmission
Engine Part Technical Trends and New Engine Products
Yoshiaki RYOUNO, Kouichi ONIMARU, Kenichi KAWABATA and Shinji OOISHI

Ball-screw Unit for Variable Valve Event and Lift System
Keisuke KAZUNODevelopment of an End-Pivot Type Mechanical Lash Adjuster
Eiji MAENO, Hiroshi BUNKO and Katsuhisa YAMAGUCHI

Transmission Technology Trends and Product Developments
Takahiro KANAMOTO, Takashi UENO, Akihiko KATAYAMA and Masanori SATOU

Dynamic Analysis for Needle Roller Bearings Under Planetary Motion
Tomoya SAKAGUCHI

Micro HL Tapered Roller Bearing
Takashi UENOHigh Speed Thrust Needle Roller Bearings
Kosuke OBAYASHI

62
72
78
86
94
100
105

● Accessory
Accessory Technology Trends and Product Developments
Ikuo FUJINIWA, Makoto MURAMATSU and Tadahisa TANAKA

Development of NA103A Long-life Grease for Automotive Components
Takayuki KAWAMURA and Hidenobu MIKAMI

Compact Clutch Integrated Pulley for Alternators
Koji SATO and Isao MIKURIYA

110
116
124

●Our Line of Award Winning ProductsSociety of Tribologists & Lubrication Engineers Captain Alfred E.Hunt Award

Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain

128

Noriyuki TSUSHIMA
JSME Young Engineers Award

Logarithmic Profile of Rollers in Roller Bearings and Optimization of the Profiles

140

Hiroki FUJIWARA and Tatsuo KAWASE
The 2005 Most Interesting Reading Award by Japan Society for Design Engineering

Run-out Analysis for Rolling Element Bearing

149

Tomoya SAKAGUCHI
Monodzukuri Manufacturing Award A prize for encouragement

Improving Rolling Contact Fatigue Life of Bearing Steels Through Grain Refinement "Development of the FA bearing"

150

Chikara OHKI
12th Advanced Display of the Year 2007(ADY) Excellent prize winning

LCD Color Filter Repair System "NRS-3000 Series"

151

Akihiro YAMANAKA

Our Line of New Products

152

NTN TECHNICAL REVIEW No.75（2007）[ Preface ]Introduction to this Special Feature Issue –About the 75th NTN Technical Review,
Which Features NTN’s Unique Eco-Conscious
Automotive Technologies
Kenji OKADA
Executive Director

The targeted reduction in greenhouse gas emissions assigned to Japan according to “the Kyoto Protocol to the United
Nations treaty on climate change” in effect in February 2005, is 6% decrease relative to the year 1990 level. Regrettably,the 2003 emission in Japan marked 8% increase over the 1990 level, rather than a decrease—many knowledgeable
people are not confident about the fulfillment of this target. At the same time, global warming has been ever worsening,
the evidences of which include damages inflicted by super-hurricanes in the U.S.; inundation in the U.K.; and an
unprecedented long spell of very hot summer in this year in Japan—various regions in Japan experienced a record-highnumber of days above -30˚C causing many Japanese to assert that “we are experiencing an abnormal summer”. The
global environment is damaged to such an extent that our daily lives are increasingly jeopardized.
Automobiles are responsible for approximately 20% of CO2 emissions in Japan (-FY2004 data). As more automobiles
are sold in the BRICs and newly burgeoning markets, the total number of cars manufactured throughout the world in2006 nearly reached 70 million. On the other hand, a series of global regulations for mitigating environmental impacts
have been proposed, and the examples of which include “Establishment of 2015 fuel economy standard in Japan” and“CO2 emissions control in Europe”. To address this situation, development work for various eco-conscious automotive
technologies are in progress and will be further accelerated. Examples of these technologies include hybrid cars, ethanolcars, diesel cars and electric vehicles.
The Japanese version of this review featuring NTN’s latest eco-conscious automotive technologies was issued,
intended for presentation in the 40th Tokyo Motor Show that began on October 26 and ended on November 11. The
main part of this review begins with a contribution titled “Saving Energy with Transmission Technology for a Motor
Vehicle” from Professor Haruo HOUJOH, TOKYO INSTITUTE OF TECHNOLOGY, who is known as an authority in the
research of automotive transmission system dynamics. Then our automobile-related technologies are presented in three
application fields—<Drive line>, <Engine & transmission> and <Electrical equipment & auxiliaries>., wherein eachsection contains information about trends in eco-conscious engineering development efforts and an introduction to
NTN’s eco-friendly products and technologies.
In this review, our new engineering developments are categorized into three engineering fields: the “electric drive
system” section presents an in-wheel motor unit and an electric brake unit; the “compact & light-weight technology”
provides a pulley assembly for alternator w/ built-in compact clutch; and the “long service life technology” section offers
knowledge about micro HL conical roller bearings and long-life grease for electrical auxiliary bearings.
The review presents various new technologies. The “sensor” section describes a technology for a high resolution
velocity & position sensor; the “analysis” section offers examples of analysis that include dynamic analysis on bearings
under planetary motion and analysis on internally working forces with constant velocity joints.
The review also includes information about NTN’s five award winning technologies for FY2006 (STLE 2006 Captain
Alfred E. Hunt Award, “Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain”; The
Award for Prospective Researcher—The Japan Society of Mechanical Engineers, “Logarithmic Profile of Rollers in Roller
Bearing and Optimization of the Profile”; The Monozukuri Buhin Taisho Shoreisho Award, “Improving Rolling Fatigue Lifeof Bearing Steels Through Grain Refinement—“Development of the FA bearing”).
NTN assigns the three years beginning with April 2007 to our new medium term business initiative “Sosei 21” which
stands for “creativeness”, “achievement” and “progress”. Based on this business concept, we, in terms of engineeringcommitments, will further enhance the value of the NTN brand by dramatically improving quality through robust design
practices; more positive proposals to our customers and further accelerating product development.
In the field of environmental conservation, we will be more positively committed to the efforts for ecology: the examplesof these efforts include more intensive development and marketing of eco-friendly products; attempts to totally ban useof environmentally impacting substances; positive utilization of clean natural energy sources; and recycling of otherwise
disposed materials.
In accordance with NTN's creed: “For New Technology Network: contributing to the international society throughcreation of new technologies and development of new products”, NTN's commitment to develop eco-friendly productsthat contribute to “coexistence with the global environment” and drive to help achieve both realization of the sustainable
society and better convenience that results from motorization.

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NTN TECHNICAL REVIEW No.75（2007）[ Contribution ]Saving Energy
with Motor Vehicle Transmission Technologies

Haruo HOUJOHTOKYO INSTITUTE OFTECHNOLOGY
Precision and Intelligence
Laboratory

Transmission technologies have been reviewed in this article from the viewpoint of
global energy conservation. First, the importance of energy conservation is emphasized
for a sustainable world not only because of global warming, but also because ofhistorical planetary equilibrium. In addition, an assessment of future automobiles in thenext 25 years is briefly introduced, predicting no promising systems among the various
solutions for reducing fuel economy. Second, the energy consumption requirements ofmotor vehicles are discussed based on Newtonian mechanics to clarify the roles of bothengines and transmissions. It follows that, although improvement of engine performanceis the most essential, transmissions also play a large role in achieving better
performance in regard to fuel consumption and drivability. Then, some key aspects for
improving transmission energy consumption are briefly reviewed. Finally, the reuse oftransmission units or parts is suggested as one effective way of saving energy.

period, various living things somehow maintained theirlife activities through a long series of rises anddeclines on the surface of the earth in the equilibrium
state. In other words, the solar energy reaching theearth is considered to be an indispensable thing tomaintain the life activities on the earth in an
equilibrium state.
An attempt to generate a plant-based hydrocarbon
product such as a bioethanol and to use this productas a car fuel may be comparable to an act of stealing
energy (figuratively, "daily income") that is needed forbasic life activities. Deforestation activities to create
farms for cultivating the sources of bioethanol alsopose a challenge to the global environment. From thisstandpoint, even wind power generation and solar
power generation may seem to help jeopardize theglobal environment. "Stealing" of energy in a very
limited area on the earth may be acceptable. We
cannot determine which factor can lead to disruptionof the earth's equilibrium. I think dependency on a few
alternative energy sources to replace a greater portionof necessary energies is not a wise option. We should
use smaller amounts of energy from more diverse
energy sources.
Therefore, in consideration of present-day energysaving efforts, we need to develop a comprehensive
energy saving measure rather than simply reducing
CO2 emissions.

1. Introduction
Many years have already elapsed since CO2-induced
global warming first posed a challenge for mankind.
Now, environmental issues have been increasingly
worsening. Energy saving is an outstanding challenge
to be addressed for automobiles that are typical
products in our consumption-oriented societies. Having
specialized in automotive drive systems, the author has
attempted to summarize the role of vehicle transmission
technology in energy saving aspects together with his
own private opinions.

1.1 The outstanding challenge is energy savings
A Japanese newspaper reported the content of thelecture that was presented, in the Agricultural
Environmental Symposium held in Tokyo on May 23,
2007, hosted by Lester R. Brown, President, the Earth
Policy Institute, USA. While use of biofuels was
encouraged to mitigate emissions of greenhousegases, he warned that the prices of food would
dramatically rise and as a result the social
environments with impoverished people around theworld would be jeopardized.
His opinion appears to be true. The fossil fuels
supporting the present-day motorization of modern
society, typically, coal and petroleum*1 are, figuratively
speaking, "bank deposits" accumulated on the earth
by solar energy in a time span as long as nearly abillion years after the birth of the earth, during which
period the earth remained in an equilibrium state as
the earth experienced several ice ages. During this

*1 Some researchers have proposed "abionic process"theories to explain the origin of petroleum. Regardless ofthe true origin, petroleum is the "bank deposit" that haslong been possessed by the earth.

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Saving Energy with Motor Vehicle Transmission Technologies

requirements of safety and eco-friendliness. In older
days, ownership of cars meant "joy of driving", and carowners enjoyed the refreshing feelings and sensations
from acceleration as well as a comfortable trip while
they drove their cars. This view of cars seems to havechanged. Most drivers of the present day seem toregard their cars only as a means of transportation,
though they hope that driving in their cars is
comfortable.
In its car industry engineering strategy for 2030
cars, JSAE set up the following six considerations as
the benchmarks: a) energy saving (2L car), b) super
low emission gas level (1/10 of the FY2005 level), c)
low noise level (1/2 noise level of the FY2001 level),d) resources saving (100% recycling), e) safety(reduction of numbers of traffic accidents andcasualties to 1/2), and f) information (media-freearrangement). Then, we issued, to knowledgeable
people, a questionnaire about technologies needed forsatisfying these benchmarks, wherein each
respondent was expected to define these benchmarksin three levels development, commercialization andcommon use. One example of a post-edited result of
the response is illustrated in Fig. 1. For more details,
the readers are encouraged to refer to the related
document. The particular information in this document,
which seems to be most closely related to the author's
contribution, is associated with efforts for energy andresource savings, and the information is hereundersummarized.
Some people expect that a 2L car (whose fuelconsumption and CO2 emissions for travel of 100 km
are equivalent to those of 2 liters of gasoline
consumed) will be commercialized and commonly

1.2 Energy is "money"
I am a layman in economics. Notwithstanding, I at
least know that production of an industrial productunavoidably consumes energy. A machining process
unavoidably involves the consumption of energy.
However, various forms of human energies which may
include a labor cost are transacted as commodities.
To sum up, the price of a product may be considered
to be an index that stands for the amount of energy
consumed during the production process for the
product in question. Accordingly, energy used to makea product can mean that the energy balance of agiven product is measured by the money spent. We
should consider that purchasing an expensive carmeans that a lot of energy has already been
consumed to manufacture this car.

2. Cars of the future
Recently, I participated in a program sponsored by
the Society of Automotive Engineers of Japan (JSAE),
in developing a car industry engineering strategy andan engineering development scenario1). In thisprogram, a questionnaire was presented to visitors in
the Tokyo Motor Show 2005, asking what things arenecessary for their daily lives. The top ranking
answers were cars, computers and mobile phones, inthat order. Similar questions were issued to university
students in Japan, and the top ranking answers were
mobile phones, computers, bicycles and then cars(the number of answers for cars was somehow
equivalent to that of motorcycles).
These days, people who wish to purchase cars
hope that their intended cars at least satisfy mostAnalysis method Example: Realization of 2L car by utilization of fuel cell15

2L car: fuel cell
Development

Previous investigation
year

2010

2015

Number of
respondents

25

46 27

5

100%

2025

2030

Realization of 2L car by adoption of fuel cell
Commercialization

Total

〜
3

Will not berealized at all.

〜

〜

〜

〜

'10 '15 '20 '25 '30

2020

Cumulative frequency graphRealization of 2L car by adoption of fuel cell
'06 '11 '16 '21 '26

30
Common use

22

14.5

2005

Frequency distribution table

Periods forrealization

22
Commercialization

76 182

Common use

80%

60%
50%
40%
20%

2025
2026
2027
2028
2029
2030

0%
2006
2007
2008
2009
2010
2011
2012
2013
2014
2015
2016
2017
2018
2019
2020
2021
2022
2023
2024

Frequency
distribution

Will not berealized at all.

Super low fuelconsumption car

Bar graphFig.1 An questionnaire example done for assessing core technologies in next 25 years.2).

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NTN TECHNICAL REVIEW No.75（2007）

3. Discussions into energy
consumption with automobiles
3.1 Thrust force and power needed by vehicles
Let us consider, based on Newtonian mechanics,
the motion of a passenger car, where the car starts,
accelerates and reaches a constant cruise speed.
Then a relationship holds between the time andphysical amounts as plotted in Fig. 2. If an
acceleration curve is given and the mass of a vehicle
is known, then the required thrust is defined and therequired power (power) can be determined. In Fig. 2,
two speed profile scenarios are coveredﾑfrom start toa cruise speed of 60 km/h and from start to a cruise
speed of 120 km/h, wherein it is assumed that thevehicle (assumed to have a displacement of 2000 cc
and a mass of 1.5 t, in accordance with theAutomotive Engineering Handbook) is subjected torolling resistance (which is not governed by thevehicle's speed) and aerodynamic resistance that is
proportional to the square of the vehicle's speed.
When the vehicle travels uphill, an additional
resistance proportional to the degree of gradient will
be added to the rolling resistance, the required thrust
will be greater, and additional power will be needed.
To cope with these physical amounts, the impedanceof the output from the engine is adjusted such that thedrive train including the transmission and final drive
provides the thrust needed for an intended vehicle
speed that also overcomes the rolling resistance in thevehicles tires.
The required thrusting force of a vehicle inacceleration is defined as a total of rolling resistanceof the tires and the vehicle's aerodynamic resistanceas well as the vehicle's "acceleration resistance"

140

7
80

Requiredthrusting force

6

Required power

Speed

120

70

50
40
30

5

100
Thin line: acceleration to
120 km/h cruise

4

80

3

60

2

40

Velocity km/h

60

Thrusting force kN

Required power kW

used by the year 2015. This epoch-making
achievement will result from a HEV (Hybrid Electric
Vehicle) design that also has a gasoline engine or an
EV (Electric Vehicle) design. Commercialization of a
FCV (Fuel Cell Vehicle) will take place in and after theyear 2020; however, many people deny the possibilityof common use of FCV even by the year 2030.
At the end of May 2007, a report about "nextgeneration vehicle and fuel initiatives", a program
mainly sponsored by The Ministry of Economy, Tradeand Industry of Japan, was issued, wherein the reportmentions that the automotive fuels will be diversified
through use of batteries, clean diesel engines,
hydrogen & fuel cells, and biofuels. As a result ofstudy into the responses from the above-mentioned
questionnaire, we find many pessimistic projections
about commercialization and common use oftechniques for production and provision of fuel cell.
This trend coincides with the intention of researches
for diversification in automotive fuels. It is expectedthat a variety of forms of so-called energy-efficient
vehicles will be developed based on various
technologies and techniques.
Incidentally, around the year 2020, commercialization
will be realized for typical technologies that support
energy saving cars, resulting from, reduction in
frictional losses and the use of super-light highstrength non-metal and metal materials for drive
systems. It appears that these particular technologies
will be commercialized relatively soon. The strategy
developed based on the findings from the abovementioned investigation stresses importance of not
only development of new power sources and further
development of low fuel consumption cars based on
the currently available technologies, but also of
associated peripheral technologies and techniques.
In addition, we hope that the ratios of recycled orreused parts in the cars of the future exceed
90%. The old practice of purchasing a used carcomponent from a junkyard can be practiced in amuch more modernized way. A more moderncar, at the end of its useful life, needs to be easily
disassembled for recycling. Furthermore, to allow
for reuse of components from used cars,
importance in improving technologies for partscompatibility design and accurate life-end
recognition has been highlighted.

20
10

1

20
Bold line: acceleration
to 60 km/h cruise

0010Time elapsed s

20

Fig.2 Typical example of thrusting force and power required during
acceleration from standstill through cruise.

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Saving Energy with Motor Vehicle Transmission Technologies

power-assist, the fuel economy of such an electric
hybrid vehicle will be much more efficient.
However, from an "energy is money" standpoint, a
car product may be regarded as an item that
consumes much energy because the initial investment(that is, vehicle price) is high. In other words, anumerical value standing for improvement in fueleconomy does not seem to directly stand for a degreeof energy saving.

which is a sum of the vehicle's mass and the momentof inertia occurring on the vehicle's rotating portions.
As to the rolling resistance of the tires and the vehicle's
aerodynamic resistance, the work from the thrusting
force is dissipated in the form of heat energy. The
work relative to the acceleration resistance isconserved as the kinetic energy of the accelerating
vehicle, and is dissipated as heat when the vehicle's
brake is applied.
The peak at around the 10 seconds point in Fig. 2
occurs from the power consumption for this conserved
energy. The power consumption corresponding withthe sum of rolling resistance and aerodynamic
resistance takes place in this period, wherein therolling resistance proportionally increases, while theaerodynamic resistance increases in proportion withthe cube of the vehicle's speed and accounts for thehighest percentage of power consumed in the cruise
mode after the 20 seconds period. Variation in therequired power depending on the vehicle's velocity
stems from the variation in aerodynamic resistance.
As illustrated in Fig. 2, a significantly large output isconsumed for acceleration, and the consumption of
output maximizes in the transition region fromacceleration to stable cruise.
This maximum value can vary greatly depending on
the pattern of transition: with "mild acceleration", therequired performance for the engine will be lessdemanding. Incidentally, a conventional engine
provides maximum acceleration in the full-throttle
setting; the engine does not develop its maximum
output in an ordinary travel situation. Also, decreasing
the above-mentioned dissipated energy will contributeto energy saving.
Conventionally, kinetic energy is stored in a vehicle
and is dissipated as heat energy when the brake isactuated. On a hybrid electric vehicle, the kinetic
energy is recovered and is used as an energy source
to drive the electric motor, in order to start the vehicle
and compensate for insufficiency in the required
output in a situation such as acceleration for passing.
With the so-called 10-15 mode, fuel economy with ahybrid electric vehicle is dramatically improved
because in this mode acceleration and deceleration
are frequently repeated and the resultant conserved
energy is reused (such conserved energy is neverreused on any conventional vehicle).
On the other hand, it is apparent that any modern
automobile usually does not require much power since
its engine runs at a lower speed of 2,000 rpm evenwhen the vehicle is traveling at 100 km/h. Therefore, if
a "less powerful, but highly efficient engine" isresponsible for cruising at a given speed and anelectric motor takes part in instantaneous positive

3.2 Operation of the engine, and energy
In the graphical plotting in Fig. 3, the engine speed
values are taken on the horizontal axis while theengine torque values are along the vertical axis,
wherein the engine performance graph is plotted withthe engine equivalent efficiency curves assuming thatthe optimal efficiency taken at 100. Also, in Fig. 3, theequivalent output curves, determined by the engine
speeds and torques, are hyperbolically plotted.
Next, suppose that the acceleration profile
illustrated in Fig. 2, is realized with an ordinary 5speed manual transmission system. Then, theoperating trend with the engine (speed and torque)can be plotted with a bold dotted line. For instance,
when a vehicle is accelerated to 60 km/h in thisdiagram, the vehicle's speed profile may consist of 1
acceleration to 2200 rpm with the 1st gearshift, 2
acceleration from 1300 rpm to 2700 rpm with the 2ndgearshift, 3 acceleration from 1800 rpm to 2800 rpm
with the 3rd gearshift, and finally 4 cruise with the 4thgearshift. As apparent from this diagram, under these
conditions, acceleration to 120 km/h means a driving
mode that is 10% or more efficient.
Incidentally, the pattern of plotting can varydepending on the smoothness in start-up of

200

Optimal fuel economy path

Equivalent fueleconomy curve

3

97%

150

Engine torque Nm

4

2

100

50

90%
80%

60k

W

5th
1st

1

1

95%

2

3

30kW

1st
4th
120km/h cruise
60km/h cruise

0
1000

2000

3000

Equivalent output curve
4000

5000

6000

Engine speed r/min

Fig.3 Engine performance of an ideal 2L gasoline engine
and operation paths-5-

NTN TECHNICAL REVIEW No.75（2007）

based on compatibility with particular diesel engines
or gasoline engines, or hybrid electric systems.

acceleration and transition to a cruise: therefore, it
should be noted that comprehensive discussions for avehicle's efficiency under fixed conditions aremeaningless. Additionally, some of the step ATs*2)
recently very often used in Japan have the lock-up
function, a vehicle's behavior immediately after
starting travel may appear somewhat different. Of
course, we need to take into account the number ofgearshifts in evaluating the efficiency of a particularvehicle.

a) Expectation for engine technology
As described above, when the intended vehicle
employs either MT or AT, the running state of theengine of this vehicle is not always situated in the
optimal efficiency region in Fig. 3. If the engine's
running state falls in the region where the arrows each
representing the state and when the efficiency of theengine is satisfactory in a wider speed range in Fig. 3,
in other words, when the intervals between thecontour lines are wider, the fuel economy of thisvehicle will be still better. Thus, we want further
improvement in the engine performance.

3.3 Expectation for drive system in contribution
to energy saving
The engineering scenario1) that was developed inconjunction with the engineering technology includes
discussions about drive systems in the future. When
we attempt to predict the prospective technologies up
to the year 2030, there appears to be a unique trend
for each nation and no one best solution seems to befound for a particular region. Various travel profiles are
possible in various nations and regions. Various
solutions are possible for various cultures. Comparedwith cars in Europe and America, the vehicles innations including Japan experience more stops andstarts. The transmission systems used can be either
AT or CVT. The drive systems most commonly usedon the cars in the European market seem to be avariety of combinations of a diesel engine and MT or
AMT (an automated MT): thus, one particulartransmission type does not appear to prevail (Fig. 4).
These transmission systems will be each selectedb) Expectation for transmission
Currently, CVTs are often used. A CVT can be
adjusted such that the optimal operation path matches
the intersections between the contour lines (efficiency)
and equivalent output curves: as a result, the vehicle
can be more efficient in terms of energy saving.
However, a CVT can pose a problem in terms oflayout (a belt-type CVT is suitable for a FF layout car
while a troidal-type CVT is intended for a FR layoutcar) as well as in torque performance. In other words,
a CVT may not be readily incorporated into any
passenger car. Because there is a possible challenge
about the efficiency resulting from a torque converter
and/or hydraulic circuit: to address this problem, there
has been an attempt to develop a clutchless CVT
system that is referred to as the IVT.
At the same time, an increased number of stages
have been applied to MT and step AT (wide range andclose ratio) so that a vehicle can be run more
frequently at an engine speed where the engine canrun at higher efficiency to achieve better fueleconomy. The effect of this effort is much apparent*2) Step AT: An automatic transmission (AT) system thatcomprises a torque converter and a stepped transmission(usually, comprising a planetary reducer group and amultiple-disk clutch).

Roadmap

HydrogenFossil fuels

Biofuels, GTL, DME

Passenger cars

Higher efficiency: Minimization of input energy
Regeneration: Recovery and reuses of input energy
1 2-motor/generator + planetary gearing mechanism
2 1-motor/generator + mechanical transmission

HEV
CVT Belt HMT Troidal

IVT

AT Increased number of gearshifts: 4, 5, 6, 7, 8…..
AMT

DCT

MT

TrucksHEV 1 1-M/G + mechanical transmission⇒22-M/G + planetary gearing mechanism
AMT

Increased number of gearshifts: 7, 9, 12, 16

MT
2000

2010

2020

2030 year

Fig.4 Sharing of various kinds of transmissions in next 25 years4)-6-

Saving Energy with Motor Vehicle Transmission Technologies

example of published achievement in efficiency
improvement for a CVT system. A new transmission
design boasts not only optimization of the system but
also "lightening, reduced drag (clutch, bearing) andreduction in other losses". It is very hard toquantitatively determine the achieved improvements.
As illustrated, contribution of each improvement isvery minor. However, the improvement in hydraulicloss is a major contributing factor. Suppose that
improvement in the hydraulic system is mandatory,
and then the next challenge is an attempt for lower
loss with individual elements.

with large trucks. A 12-stage transmission has beencommercialized, and transmission of more increased
stages will be marketed soon.
As discussed above, people are attempting to selectan optimally performing drive system from a diversityof varieties. For this purpose, it is still necessary to
make steady efforts for realizing both good fueleconomy and running performance. Nevertheless, wehope that the automotive engines have more
improved performance since improvement in engine
efficiency helps realize further improvement in fueleconomy as mentioned in a) above.

c) More efficient bearings-From test calculation for loss on bearing for
hypoid pinion-

4. Energy saving and mechanicalelements

A hypoid gear is of course used on FR layout cars.
In particular, two hypoid gears are usually used on
4WD layout cars. A tapered roller bearing used for
hypoid drive pinion is often used with a greater thrust
preload in order to fix the attitude of the shaft, and as
a result, a greater dragging resistance occurs. Let us
determine the approximate loss under the following
conditions:Tire speed:
10 r/s (63 rad/s, approximate vehicle speed;
72 km/h)Reduction ratio with hypoid gear: 3
Dragging torque on bearing*3: 1 Nm by 2 pcs.
From a calculation under these conditions:Power loss from dragging ≒ 0.4 kW
Assuming that the maximum output of the engine issomewhat lower than 100 kW, the loss does not

For Figs. 2 and 3, no discussions were made about
the efficiency of a drive system including atransmission. Currently, the transmission efficiency
with MT appears to exceed 95%. On the other hand,
the transmission efficiency with AT or CVT seems tofall in a range of 80% to 90%, though the efficiency
can vary depending on the vehicle's running
conditions. The factors that seem to be responsible forthis variation in transmission efficiency are use of atorque converter and a hydraulic pump and greaterdrag on a multiple disk clutch. Gradual improvement
will reflect these factors.
Elements including hypoid gears are indispensable
for automobiles; however, some of these units stilllead to a greater power loss. The efforts for improved
efficiency are in progress, and some examples ofthese efforts are described below.

a) Lighter weight
When thinking of a simple dynamics, motion energyof a vehicle can be recovered by a generator;
therefore, it may seem that the vehicle can be heavy.
However, the battery on any vehicle has limitation inits capacity and the rolling resistance on the tires isusually proportional to the vehicle weight. Therefore, alighter vehicle weight helps realize better fueleconomy. Of course, energy regenerationaccompanies a loss. Therefore, lowering the vehicle
weight is very important, and achievable byimprovement in materials, sophistication in individual
elements and integration of elements.

Improvement of approx. 10%

Oil
pump

FWD/REVsystemBELT&PULLb) Reduction in agitation loss and pump loss
The typical causes that are responsible for a larger
portion of loss on a transmission and reducer areagitation loss with the lubricant and energy
consumption by the hydraulic circuit. Investigation into
the amount of such a loss has been of course
available; however, the results of quantitative studies
have been seldom published. Fig. 5 illustrates one

Oil
pump

FWD/REVsystemBELT&PULL

Differential system

Differential system

OthersOthersCurrent CVT3

Improved CVT3

Fig.5 Loss reduction achievement by revising potential
causes on a CVT unit (NISSAN) 3)*3) Similar values are included in the reports from various
bearing manufacturers, and this value has been set as a
typical value.

-7-

NTN TECHNICAL REVIEW No.75（2007）

appear to be too great.
Incidentally, the power required (determined as
illustrated in Fig. 2) to compensate for the runningresistance is approximately 8 kW. Within this power
requirement, the loss from rolling resistance of thetires is as low as approximately 3 kW when the rolling
resistance per tire is assumed to be approximately 35
N. At the same time, since the efficiency of a MT
system is 95% and that of an AT system isapproximately 85%, it is estimated that a loss ofapproximately 1 kW occurs under these running
conditions. Considering these numerical values, theloss occurring on the bearing should not be neglected.
To sum up, I hope that in the future we can publicly
discuss the amount of power loss from dragging rather
than the reduction in the dragging torque.

Power Split Device
Motor Speed Reduction Deviced) Motor/generator system and mechanicalelements

Compound Gear

Fig. 6 provides a cross-sectional view of a
motor/transmission-reducer unit used on a TOYOTA's
hybrid electric vehicle, where the maximum speed ofthe motor exceeds 12,000 rpm. This very high runningspeed with the electric motor has resulted from efforts
for more compact size and higher motor speed toachieve a lighter power package since it was possible
to assume that the motor output increases inproportion with the motor speed. When the vehicle
starts to run and accelerates, the tire speed is very
low and falls in a range of 100-200 rpm therefore, thetransmission needs to have a reduction ratio of
approximately 1/100. At the same time, so that thishybrid electric vehicle system can function as aparallel hybrid electric vehicle system, a power split
mechanism involving a planetary gear mechanism isincorporated in order to allow the engine and the
motor/generator to develop power either independentof each other or in cooperation, wherein the planetary
gear mechanism employs a rolling bearing.
The transmission shown in Fig. 6 seems to becompact and light-weight. Since this transmission
employs a larger rolling bearing that supports the ring
gear, considerations about its efficiency are needed.
Naturally, the planetary gear (pinion) and the needle
roller bearing that supports the planetary gear appear
to reflect various considerations and the planetary
gear itself has been accordingly optimized. Though I
don't know how high the level of the entire mechanicalefficiency of this transmission is, I, as a specialist inmechanical elements think there will be various
possibilities in further improving such a transmission
system.

Fig.6 Cross sectional drawing of an up to datetransmission for a hybrid car (TOYOTA) 5)

a lubricant can help significantly improve efficiency ofa transmission (currently, a near-impossible
challenge). Of course, since contact-induced stressoccurring inside a rolling bearing will deform thebearing, and this deformation lead to occurrence ofhysteresis loss; as a result, complete avoidance ofheat generation seems to be impossible.

5. Expectation for recycling or reuse
A drive system plays a vital role in transmitting theoutput from an engine to the road surface as athrusting force. Currently, we seldom experience a
failed transmission on a passenger car. The effective
service life of a transmission for a passenger car will
not be reached as long as the car is operated in anormal manner. In a transmission for a truck, thegears for low gearshift used for starting, accelerationand traveling uphill seem to be designed not to fail,through a finite life design technique that is based onrun frequency data.
While an ordinary passenger car having traveled inexcess of 100,000 km is often scrapped, a trucke) Expectation for oilless rolling bearing
Lubrication-induced loss is present on any rolling
bearings. Therefore, running a rolling bearing without-8-

Saving Energy with Motor Vehicle Transmission Technologies

having traveled more than 1,000,000 km with its fullload capacity will continue to be used. However, inmost cases, the transmission on such an aged truck
rarely fails. Consequently, I expect that the following
practices may be being observed:

6. Conclusion
From the viewpoint that energy saving is an issue tobe addressed immediately, the author has discussed
the present-day automotive drive systemtechnologies. An automotive drive system develops its
design role as its many mechanical elements work incooperation with each other. I, as a specialist inmechanical elements, believe that we should remaincommitted to the permanent engineering challenge ofreducing heat occurring on a mechanical element tonear-zero. Also, we must develop a technology that
addresses energy consumption in terms of global,
long-term viewpoint, rather than only considering thelifecycle of each part in a machine element. While
reconsidering a previous practice of "produce-andscrap" with various industrial products including
automobiles, I believer engineering people in Japan
will further promote their research and development
works, aiming at further evolution of their specialty
element technologies. I hope that NTN's engineers
and technicians will succeed in these challenges.

a) Expectation for reuse
A few lucky still operable units may be recycled
while most of still operable units are simply scrapped.
This is a great waste. If still sound units aretransferred to other vehicles, great energy savings canresult. Such a practice, even though certain parts andcomponents need to be replaced with new ones, will
mean much mitigated environmental impacts.
As already described in Sec. 3.3, if thecharacteristic curve of a particular engine is flat, thecar incorporating this engine will operate at a higher
efficiency without a procedure for fine-adjustment ofgears. This is true with a vehicle equipped with a CVT.
Though a lighter vehicle weight is generally
preferable, a certain increase in a vehicle weight may
not adversely affect the fuel economy of the vehicle. If
this assumption is true, transmission manufacturers inJapan need not manufacture more transmission types;
instead, they may produce a limited number ofstandard transmission types. This arrangement will
help a used transmission be transferred to anothervehicle.

References
1) Society of Automotive Engineers of Japan, Inc.:
Automobile technology strategy and technology
development, Fuel scenario, May 2007.
2) Society of Automotive Engineers of Japan 2007 Spring
Conference, Automobile Production Technology
Forum Text, p.15.
3) Reference 1 above, part 2, p. 34.
4) Seiji Terauchi, Minoru Sawayama, Tatsuo Ochiai,
Akihiro Makiyama, Masanori Ishido: Development of
New Xtronic CVT for High-torque capacity,
Symposium text of Society of Automotive Engineers ofJapan, No.15-06 pp.18-22 (2006)
5) Hiroshi Hata, Masahiro Kojima, Hideto Watanabe,
Tatsuhiko Mizutani, Munehiro Kiyama, Kenji Takizawa:
Development of a Hybrid Transmission Engineers ofJapan, No.10-05 pp.47-52 (2005)b) Improved reliability for life design
For present-day machine elements, failure of amachine element due to overloading is a simple
problem in dynamics design; however, a more
important problem is fatigue failure. The mode of
fatigue failure is said to be governed not only by thecomposition and crystal structure of the material of thefailed machine element but also by the inclusions in
the material. Therefore, even now, fatigue failureneeds to be discussed stochastically. In theengineering field of rolling bearings, the concept of life
design has been accurately developed.
Notwithstanding, we can predict the remaining life of acurrently used part to an accuracy of "a specificnumber of days left" only when the operating
conditions of that part are strictly limited.
The concept of finite life design appears to havebeen commonly accepted. I believe that further
sophistication of this technique will help reduce wasteof energy through recycling and reuse of machineparts.

〈About the author〉
Haruo HOUJOHProfessor, Tokyo Institute of Technology Precision
and Intelligence Laboratory
Field of expertise: Dynamics of machine elements(especially gears)
Academic organization memberships:
Japan Society of Mechanical Engineers,
Society of Automotive Engineers of Japan-9-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Technical Trends in Constant Velocity Universal Joints
and the Development of Related Products

Shin TOMOGAMI*
Constant Velocity Universal Joints (CVJ) have come to
be used widely for driveshafts and propeller shafts invehicles, and are now being used for almost 100% ofJapanese cars. CVJ performance has improved greatly
with the evolution of vehicles. This paper will introduce
recent technical trends and the development of products(parts) related to driveshaft CVJs.

This paper describes the technical trend about the
recent driveshaft CVJs as well as newly developed
CVJ products.

1. Introduction
Forty-four years have already passed since NTN
began production of the Constant Velocity Joint(hereinafter simply referred to as "CVJ"). The total
number of CVJs produced thus far for driveshaft has
nearly reached 400 million.
As Japan's auto industry has been expanding
throughout this period, CVJs have been commonlyused on driveshafts and propeller shafts of FF (Frontengine Front-drive) layout cars, IRS (Independent
Rear Suspension) layout cars and 4WD (Four-Wheel
Drive) layout cars. In today's market, virtually every
passenger car designed in Japan incorporates the useof CVJs. The typical vehicle related areas in which the
CVJs are applied have been illustrated in Figs. 1 and
2.
As automotive designs continually evolve, theperformance of CVJs has been improved accordingly.
FF layout

IRS layout

2. Current requirements for CVJs
Around the world, the global environmental issues(including global warming) have been posing
challenges that need to be immediately addressed.
Automobile manufacturers and their suppliers havebeen committed to various efforts including improved
fuel economy, introduction of hybrid car design, use of
bio-fuel and development of high-performance diesel
engines in order to reduce emissions of CO2 which isa major contributing factor to global warming.
As a result of these initiatives, CVJs as drive trainrelated components are required to feature a lighter
weight and offer better torque transmission efficiency
in order to help improve fuel economy on the vehicle.

4WD layout

CVJ

CVJ
CVJ

Ball-fixed
CVJ

Fig. 1 CVJ application*Automotive Sales Headquarters

Plunging
CVJ

Fig. 2 Example: drive shaft of front axle

C.V.Joint Engineering Dept.

-10-Technical Trends in Constant Velocity Universal Joints and the Development of Related Products

In the next section, the author will present thehistory regarding weight reduction of driveshaft CVJs,
Thereafter, the author will further explain functions,requirements and recent achievements about ballfixed CVJ, plunging CVJs and CVJ components; eachbeing a constituent of a driveshaft.

4. Ball-fixed CVJs
4.1 Functions of ball-fixed CVJs
A ball-fixed CVJ used on a front driveshaft of an FF
or 4WD layout car is coupled to the front wheel hubs
on a vehicle. This application results in a working
angle while the steerable tires change the running
direction as the steering wheel is operated, thereby
transmitting the driving power from the engine to thetires while smoothly running at a constant velocity.
Additionally, a ball-fixed CVJ directly affects theunsprung weight of a vehicle suspension.

3. History of weight reduction fordriveshaft CVJs
As the Holling-Kelly CAFE proposal seemed to be
imposed in the US, the efforts for lighter weightcomponent designs were strongly needed for every
automotive manufacturer since the latter half of the
1980's. To cope with this need, NTN has beencommitted to light-weight, compact designs for its
driveshaft CVJs with its unique technologies.
In the early 1990's, NTN developed a high-strength
material and a long-life lubricant (grease) to realize alighter-than-ever driveshaft CVJ. As a result of thiseffort, a lighter weight arrangement was achieved,
whereas a CVJ of one nominal size smaller relative toa CVJ previously used on a given vehicle size could
be employed.
In a time span from 1998 to the earlier half of 2000,
NTN developed and mass-produced the nextgeneration CVJ products (E-series CVJ) that aremuch more compact compared with the previous CVJ
series products, in order to meet the needs from theautomakers. The change in size and weight with atypical CVJ (87 size) is graphically plotted in Fig. 3.
Compared with weight levels in 1990, the weight ofthe present-day ball-fixed CVJ is 23% lighter whilethat of the present-day plunging CVJ is 17% lighter.

BJ87

BJ82
3.7

OD
(mm)
90

A need for a smaller turning radius with a vehicle,
that is, a need for a greater steering angle with front
wheels, has been significantly increasing as the recent
car designs involve longer wheel bases and morediversified drive train layouts. In this context, a greater
working angle is needed for ball-fixed CVJs whose
working angles are governed by the steering angle offront wheels. Previously, the maximum working angles
needed for ball-fixed CVJ were at approximately 47
degrees. However, larger working angles with ballfixed CVJs on many car designs are currently neededwith a (maximum working angle of up to 50 degrees).
At the same time, decreasing the weight of the car
is necessary in order to help improve fuel economy,
and under this circumstance, lighter CVJs are strongly
needed. In particular, because of being situated near
the tires, ball-fixed CVJs contribute to the unsprungweight of a vehicle. Therefore, a decrease in theweight of these CVJs positively contributes to adecrease in the overall vehicle weight.

EBJ82

TJ87

TJ82

7.3

4.1

φ83.6
(OD)

ETJ82
5.6

φ78.7
(OD)

φ79.9

80

4.2 Needs for ball-fixed CVJs

Weight
(kg)

φ74.6

φ72.6

φ69

70
1.66
(Weight)

1.52

1.54

60

1.4

1.28
50
year

2.0

1.68
(Weight)
1.5

1.0
1992

1998

Present
levels1992

2002

Present
levelsFig. 3 Transition of outer diameter and weight of CVJ (Example of initial 87 size)-11-

NTN TECHNICAL REVIEW No.75（2007）

Generally, it is said that the effect by a decrease inunsprung weight is approximately 10 times effective
as a decrease in sprung weight. Thus, lighter CVJs
will greatly help improve fuel economy of the vehicles.

NTN additionally offers a lineup of the "EBJ series"
with a maximum working angle of 47 degrees
intended for vehicles on which the maximum working
angle needed is less than 47 degrees. Compared withthe "EUJ", the EBJ ball-fixed CVJ products feature a
3% smaller outside diameter and a 7% lighter weight
while maintaining the high efficiency and performanceof the EUJ series CVJs.
An automobile manufacturer can choose to use
either an EUJ or EBJ product according to the
maximum working angle needed by an intended
vehicle design.

4.3 Latest ball-fixed CVJs
As the needs from car designs continue to evolve,
the typical ball-fixed CVJ products that are currently
used on the front driveshafts of FF cars and 4WD cars
are matched with reduced size (compact) and lightweight joints with a maximum working angle of 50
degrees.
NTN offers a full lineup of the "EUJ series" compact,
light-weight CVJs with a maximum working angle of 50
degrees for cars ranging from small-displacement
engines to that of larger SUVs (Sports Utility Vehicles).
Compared with conventional 50-degrees working
angle capable CVJ designs, the "EUJ" series products
boast approximately 7% reduction in the outside
diameter and approximately 15% reduction in theweight (Photo 1).
Furthermore, the EUJ series products are lightweight, high-efficiency eco-friendly CVJs products thatfeature 25% reduction (working angle of 6 degrees) in
the loss in driving force (torque loss) transmitted froman engine to the tires (Fig. 4).

4.4 Vibration reducing techniques for ball-fixed
CVJs
It is said that a plunging CVJ in a driveshaft affects
the NVH (Noise, Vibration and Harshness)
performance of a vehicle. Additionally throughresearch, it has been determined that thecharacteristics of a ball-fixed CVJ also affects the
NVH performance of a vehicle.
In particular, reducing the bending load (resistanceoccurring when a CVJ takes a working angle) on aball-fixed CVJ can dampen the idling vibration on thevehicle. "Idling vibration" can be defined as minutevibration that is transmitted from the running engine tothe vehicle body through the CVJ when the footbrakeon an automatic transmission vehicle (A/T car) at a
temporary stop is applied while the transmission is in adrive gear.
NTN achieves reduction in the bending load of its
ball-fixed CVJs through optimization of the internal
component design and dimensions (Fig. 5, Photo 2).

Reduction in bending
load by optimization ofinternal design &dimensions

Low ← Torque loss ratio → High

Photo 1 Compact & light weight fixed CVJ with 50
degree working angle (EUJ)

Reduction in
idling vibration

14%
improvement

Conventional CVJ (UJ)
Compact & light-weightball-fixed CVJ (EUJ)

Fig. 5 Method of idling vibration reduction

25%
improvement

0

10

20

30

Working angle (deg)Fig. 4 Measurement data of torque loss forcompact & light weight fixed CVJ (EUJ)

Photo 2 Less idling vibration CVJ-12-

Technical Trends in Constant Velocity Universal Joints and the Development of Related Products

coupled with the differential. Moreover, the roller
cassettes can stably roll on the rolling surface of theouter ring when the CVJ develops a working angle
and/or axially plunges. This arrangement helps reduce
the plunging resistance and internal friction so as toachieve ultra low vibration (Fig. 6).
Our ultra-low vibration CVJ products excel intransmission efficiency for driving force (torque). When
combined with the previously mentioned EUJ and
EBJ, the PTJ will help minimize the torque loss on thedriveshaft into which it is incorporated.

5. Plunging CVJs
5.1 Functions of plunging CVJs
Being coupled with a differential, a plunging CVJ
transmits the driving power from an engine and atransmission to tires via a ball-fixed CVJ. A plunging
CVJ is capable of absorbing the change in position
relative to a ball-fixed CVJ that results from move of
the suspension (axial displacement and angulardisplacement).

5.2 Requirements for plunging CVJs
5.4 Ultra-low vibration CVJ of compact & lightweight design

It is widely known that the characteristics of any
plunging CVJ affect the NVH characteristics of a given
vehicle. Axial plunging resistance, which occurs whena plunging CVJ vibrates in the axial direction, is
responsible for idling vibration of the vehicle. In
addition, the axial force (known as "induced thrust")
occurs on a plunging CVJ when the CVJ takes aworking angle and runs while being subjected to adriving force. This force can contribute to a jerking
phenomenon when the vehicle starts to travel as wellas produce howling sound while the vehicle is runningat a higher speed.
Therefore, plunging CVJs are required to havelower plunging resistance lower induced thrust andincorporate weight reduction design characteristics

To satisfy the market needs, NTN hascommercialized compact, light-weight ultra-low
vibration "EPTJ" series CVJ products that feature astructure and low vibration characteristics identical tothat of the ultra-low vibration "PTJ" series CVJs.
Further compared with the PTJ, the EPTJ boasts
approximately 4% reduction in outside diameter andapproximately 8% reduction in the weight. NTN offers
the EPTJ series products for larger CVJs (NTN
nominal size 95 or greater) with which a lighter weightprovides many advantages (Photo 3).

5.3 Latest plunging CVJs
To cope with recent needs for improved quietness
and riding comfort with vehicles, CVJ manufacturers
are developing a diversity of plunging CVJ designs.
Having developed an ultra low vibration plunging CVJ"PTJ" that boasts greatly reduced plunging resistanceand induced thrust, NTN products contribute toimproved NHV performance of vehicles.
Our "PTJ" is a tripod type plunging CVJ design,wherein three roller cassettes (each being a special
roller bearing) are situated inside the outer ring that isRoller cassette

Trunnion

Photo 3 Ultra low vibration CVJ with compact &lightweight design (EPTJ)

Outer ring
Rolling surface on outer ring

[Mechanism of ultra-low
vibration arrangement]
The roller cassettes are held parallelwith the rotation axis and stably rollon the rolling surface of the outer ring,
thereby the friction force between
parts is minimized so as to achieveultra-low vibration.

Rotation axis
θWorkingangle

Fig. 6 Structure of ultra low vibration plunging CVJ (PTJ)-13-

NTN TECHNICAL REVIEW No.75（2007）

Therefore, in order to realize both lighter weight andhigher rigidity, a shaft being hollow over its entire
length may be used (Fig. 8).

6. CVJ components
6.1 Requirements for CVJ components

6.3.1 Features of hollow shaft

The requirements for CVJ components relative tothe vehicle include lighter weights and enhanced
rigidity (stiffness) against torsion on driveshafts.
Greater rigidity on a drive system with CVJs will
positively improve driver's feeling while maneuvering
the vehicle. Therefore, highly rigid CVJ components
are strongly needed especially for sporty and/or luxury
passenger vehicles.

A length of steel pipe such as a one shown in Fig. 9
is swaged with a special machine (metal forming) tocontrol the outside and inside diameters of the shaft
while reducing the overall shaft diameter, whereby ahollow shaft is then formed.
Next, the hollow shaft is heat-treated (induction
hardening, etc.) to increase its mechanical strength.
Upon completion of this unique design process, thehollow shaft can be up to 20 to 30% lighter comparedwith a conventional solid bar shaft whose torsional
rigidity is equivalent. Additionally, a hollow shaft
design also boasts increased bending rigidity, andthus can dampen the vibration resulting from bending
resonance on the shaft.

6.2 Latest technology with CVJ components-hollow design
A lighter, more compact component usually has
lower rigidity. This is true with the components of
CVJs too. In particular, rigidity is greatly affected by
the shape of a shaft that connects a ball-fixed CVJ toa plunging CVJ. A recently available technology thatsatisfies conflicting requirements of lighter weight andhigher rigidity is a hollow shaft design. A few
examples of this technology are introduced below.

6.4 Hollow design--application to plunging CVJ
With a transverse engine layout such as those on
FF layout cars, the differential is offset relative to thecenterline of the car. Consequently, the length of theright half of the driveshaft differs from that of the left
half. With certain designs of driveshafts used on thistype of car layout, the right and left stems to beinstalled to the differential of the plunging CVJ eachhave a unique length so that the length of the right halfof the driveshaft is same as that of the left half of thedriveshaft (equal shaft length design). With thisarrangement, the working angle of the ball-fixed CVJ
is same with the plunging CVJ on each of the right halfand left half of driveshaft. As a result, the vehicle
boasts better driving stability; therefore, an equal shaft
length design is adopted for many FF and FF based
4WD vehicles.
With an equal shaft length design, the stem on the

6.3 Hollow design--application to shaft
A shaft links a ball-fixed CVJ with a plunging CVJ totransmit engine torque to the wheels. Conventionally,a solid bar made of carbon steel for machinery is heattreated and is used as a shaft (Fig. 7). Compared with
the CVJs (both ball-fixed and plunging CVJs), theshaft typically has lower torsional rigidity. Therefore, it
is an effective arrangement to increase the rigidity ofthe shaft in order to enhance the torsional rigidity ofthe entire driveshaft.
Reiterating that a larger diameter leads to increased
rigidity of the shaft, one must also note that thisarrangement in turn results in a much greater shaft
weight.

Fig. 7 Solid bar shaft (Conventional design)

Fig. 8 Hollow shaft
Material (steel pipe)(Swaging + partial turning)
Post-machining shape
Heat treatment
Fig. 9 Example of hollow shaft production process-14-Technical Trends in Constant Velocity Universal Joints and the Development of Related Products

The length of this section on the right-hand driveshaft differs from that on the left-hand driveshaft.

Ball-fixed CVJ

Shaft

Plunging CVJ

Stem (solid bar)

Fig. 10 Example of conventional drive shaft (Long stem design for plunging CVJ)
JointsCVJ outer ring

Steel pipe
Hollow section

Stem

Fig. 11 Example of plunging CVJ with hollow type stem

right or left plunging CVJ is 20 to 30 cm longer
compared with the other plunging CVJ and is of thesolid stem type. If sufficient torsional rigidity is to beprovided for a given driveshaft, the diameter on thislonger stem section needs to be larger, and theresultant increased weight of this section leads to aheavier overall weight with the entire driveshaft
assembly (Fig. 10).

of the previous wet grinding process that uses agrinding wheel.
The previous wet grinding process required a large
amount of coolant (oil-based and/or water-soluble
cutting fluid), which must be renewed andappropriately disposed at regular intervals. The dry
cutting process does not need any such coolant.
These are just some of the examples regarding NTN's
commitment to reduction in environmental impacts.

6.4.1 Plunging CVJs with hollow stem
According to our new technology, a length of steel
pipe is incorporated as an intermediate section for thestem between the plunging CVJ main body (outerring) and the differential in order to realize an
integrated structure (Fig. 11). The mechanicalstrength of this steel pipe section is increased by
optimizing its composition and machining method; theoutside diameters of the steel pipe and joints are
sufficiently large and the steel pipe is not heat-treated.
With the shown example hollow driveshaft, the stem
(having torsional rigidity same as with the solid
plunging CVJ outer ring) is 0.8 kg (approximately
17%) lighter compared with the solid design.

8. Afterword
This paper has presented some examples of NTN'slatest technology for driveshaft CVJs. Performance ofautomobiles improving. Functions and performance ofconstant velocity joints are directly or indirectly
contributing to this performance improvement. As a
specialist in constant velocity joint production, NTN iscommitted to developing and supplying constantvelocity joint products that contribute to the progressof automobiles and are friendly to the global
environment.

7. Eco-friendly CVJs

References

Car manufacturers are committed to design andmanufacture of "eco-friendly cars". In this context, wehave been making efforts not only to design compact,
light-weight and highly efficient CVJs that contribute tobetter fuel economy but also to eliminate any
environmentally unfriendly chemicals or substancesfrom the materials used to make our CVJ products.
In terms of eliminating environmentally unfriendly
chemicals and substances, we have alreadyeliminated grease containing lead, and paint with leadfree materials, and have adopted hexavalentchromium-free plated components instead of previous
hexavalent chromium-plated components.
Additionally, in the latest CVJ production process at
NTN, we employ a unique dry cutting process instead1) Tomoue: Latest technology trends for constant velocity
joints, Tribology Monthly, No.218, pp.49-51 (2005)
2) Takeshi Ikeda: History of CVJ Design for Automobiles
and Recent Technology, NTN Technical Review,
No.70 (2002)

Photo of author

Shin TOMOGAMI
C.V. Joint Engineering Dept.
Automotive Sales Headquarters-15-

NTN TECHNICAL REVIEW No.75（2007）[ New Product ]Fixed Constant Velocity Joint
with a Super High Operating Angle of 54 Degrees (TUJ)

Manabu HOSHINO*
Masashi FUNAHASHI*NTN Corporation has succeeded in developing a fixed-type
constant velocity joint called TUJ that has the world’s highest
maximum operating angle --54 degrees--for automobile drive
shafts. TUJ will allow four-wheel drive and front-wheel drivecars to have a greater steering angle, enabling a very small
turning radius. This report summarizes the concept, design andcharacteristics of TUJ.

1. Introduction

2. Driveshaft defined

Recently, the needs from car users have been
diversifying, and a greater variety of car designs havebeen marketed. In this context, the car manufacturers
are attempting to market unique products by adding
unique additional features.
Examples of such additional features may include
sharp turning capability and increased passenger
space. To satisfy these requirements, a vehicle mustfeature a smaller turning radius and an extended
wheelbase. These requirements can be satisfied by agreater steering angle with the tires. This particularneed has been strong for both FWD and 4WD
vehicles where steering wheels also function as
driving wheels. The tire steering angle is governed by
the design limitations as well as the maximum bending
angle (hereinafter referred to as "maximum operating
angle") of the ball-fixed constant velocity universal
joint; hereinafter referred to as "CVJ") in the tire side
within the front driveshaft system that transmits theengine output from the differential gear to thedriveshaft.
To address this need, we have recently developed asuper high angle fixed CVJ "TUJ" (Tapered-track
Undercut-free Joint) series capable of greater
operating angle compared with conventional fixed
CVJs in order to enhance the additional value andfreedom of design on the vehicle side.

A driveshaft is a power transmitting component thatconnects the engine power (revolutions and torque) tothe tires. It remains capable of running at a constant
speed and smoothly transmitting the torque evenwhen an angle occurs between the running input shaft(differential gear shaft) and the rotating output shaft(tire shafts). Generally, a driveshaft assembly in eachof the right and left sides comprises a shaft that links afixed CVJ with a plunging CVJ, wherein the fixed CVJ
cannot slide in the axial direction though being
capable of a greater operating angle while theplunging CVJ can slide in the axial direction though
being capable of only a limited operating angle. Morespecifically, the fixed CVJs are situated in the tiresides while the plunging CVJs are located near thedifferential gear (see Fig. 1).

*Automotive Sales Headquarters

Shaft
Plunging CVJ
Differential gear

Tire
Fixed CVJ

Fig. 1 Example of front drive shaft

C.V.Joint Engineering Dept.

-16-

Fixed Constant Velocity Joint with a Super High Operating Angle of 54 Degrees (TUJ)

balls in the ball track is greater with a greater
operating angle.
Therefore, to realize a greater operating angle, agreater axial length of the ball tracks is necessary.
The range of axial travel of the balls at a higher angle
situation is illustrated below, using the EBJ as an
example (see Fig. 3).

The maximum operating angle with the NTNﾕscurrently mass-produced EBJ (High Efficiency Ball
fixed Joint) products, which feature light weight,
compact size and high efficiency, is 47 degrees, while
the maximum operating angle with the EUJ (High
Efficiency Undercut-free Joint) products is 50 degrees.

3. Super high angle fixed CVJ "TUJ"

Range of axial travel of balls at high angle situation

Compared with the conventional high angle fixed
CVJs (maximum operating angle of 50 degrees), the
TUJ series CVJ products boast a greater maximum
operating angle. The TUJ series CVJs are described
below.

Ball
Cage

3.1 Features of TUJBall trackon outer ring

The components of the TUJ are identical to thoseon conventional fixed CVJs, and are composed ofouter ring and inner ring each having ball raceway
track (hereinafter referred to as ﾒball trackﾓ, six balls
seated in the ball tracks and the cage that holds theballs (see Fig. 2).
The typical features of the TUJ are as follows:[1] Maximum operating angle: 54 degrees[2] Functions (strength and durability): equivalent
to those of conventional fixed CVJs
[3] Transmission of torque is possible even at an
operating angle of 54 degrees.

Outer ring
Inner ring
Ball trackon outer ring

Fig. 3 Construction of EBJ at high-angle

The ball tracks on the EBJ are circular arc-shaped,
and the maximum operating angle of the EBJ is 47
degrees. In contrast, the ball tracks on the EUJ (highangle CVJ) are arc-shaped + straight-shaped, thereby
a sufficient ball track length in the axial direction isrealized and the maximum operating angle of 50
degrees is achieved. The newly developed TUJ, whichis capable of an extremely high range of motion, has aunique ball track shape (arc + tapered) in order toincorporate the longer ball tracks (see Fig. 4).
[2] Adoption of offset cage (see Figs. 5 and 6)
The ball track shape adopted for the TUJ (arc +taper) causes the ball tracks on the inner ring andouter ring to force the balls toward the outer ring
opening side when the TUJ is at a greater operating
angle; consequently, a problem of insufficient strength
occurs with the cage that holds the balls.
On fixed ball CVJs, due to of the nature of theirmechanism, an offset for controlling the balls isnecessary (offsetting of the ball track centers on theouter and inner rings to the axial direction);
accordingly, the ball tracks are offset relative to thecenter of sphere (hereinafter referred to as "ball track
offset"). However, this ball track offset method is
unfavorable due to structural limitations, the depth ofball track at the far side on the outer ring is shallower
when the CVJ takes a greater operating angle.
The TUJ employs an offset cage (a cage with which
the centers of the inner and outer spheres are axially

Cage

Ball

Inner ring

Outer ring

Shaft

Fig. 2 Composition of TUJ

3.2 Basic structure of TUJ
In order to be able to transmit torque at a super highoperating angle in excess of 50 degrees (anapplication not possible with conventional fixed CVJs),
the TUJ incorporates NTN's unique design
arrangements. The most unique points of this design
are described below:[1] Tapered ball track (patent pending)
On a fixed CVJ, the amount of axial travel of the-17-

NTN TECHNICAL REVIEW No.75（2007）

Thus, the TUJ design can solve not only insufficient
joint strength at high angles situation but also possible
problems resulting from an insufficient depth of balltracks.

offset) to maintain a necessary offset by means of acage, wherein the thickness of cage can be larger atthe outer ring opening side where a greater cagestrength is needed and, at the same time, possible
decrease in the ball track depth can be prevented.

Arc + taper shape

Arc + straight shape

Arc shape

EBJ

EUJ

TUJ

(max. operating angle of 47 degrees)(max. operating angle of 50 degrees)(max. operating angle of 54 degrees)Fig. 4 Comparison of a ball track contour of fixed CVJs high-angle

Centerline on balls
Center to cage
Center to cage
inner sphereouter sphere(center to inner(center to outerring outer sphere)
ring inner sphere)

Center common to outer and inner ringsand cage spherical surface
Center to inner ring ball track

Center to outer ring ball track

Offset on cage
spherical surface
Offset on
inner ring

Offset onouter ring

Typical application of offset on ball tracks(cross-sectional view on EBJ)

Typical application of offset cage
(cross-sectional view on TUJ)

Fig. 5 Comparison of structure between ball track offset and cage offset

Uniform track depth
Force for pushing
the balls outward
Track depth will be shallower withthe conventional designs.
(The illustration shows the EBJ)
Increased rigidity for the cage
(opening side) by adoption of
an offset cage design

Fig. 6 Construction of TUJ at high angle-18-

Fixed Constant Velocity Joint with a Super High Operating Angle of 54 Degrees (TUJ)

and dynamic mechanism analysis (see Fig. 8),
thereby we have realized the mechanical strength at
an operating angle of 54 degrees with the TUJ that isequivalent to the mechanical strength at an operating
angle of 50 degrees with conventional fixed CVJs.
Fig. 9 provides comparison in terms of dynamic
torsion strength.

4. Joint functions
The TUJ features the performance comparable tothat of conventional high angle fixed CVJs (maximum
operating angle of 50 degrees). In consideration of anincreased operating angle, we set the strength of the
TUJ at a maximum operating angle equal to that of theconventional high angle fixed CVJs. More specifically,
we have adopted the above-mentioned offset cage
and fully utilized an FEM stress analysis (see Fig. 7)

5. Conclusion
By employing NTN's unique design, we havedeveloped an innovative fixed CVJ featuring amaximum operating angle of 54 degrees while having
the performance comparable to that of conventional
CVJs (maximum operating angle of 50 degrees). Thismaximum operating angle (54 degrees) for automotive
driveshaft is currently the highest in the automotive
industry. The increase of 4 degrees in operating angle
means that the minimum turning radius of an averagemedium-sized FF car can be decreased byapproximately 70 cm (13%) or the wheelbase of suchclass of a car can be enlarged by approximately 40
cm (15%) while maintaining a minimum turning radius.
In other words, a medium-sized car can feature aturning radius of a compact car or the passenger
space of a larger car allowing designers more
flexibility in consideration of layout and design.

Fig. 7 FEM stress analysis for cage

References
1) Journal of Society of Automotive Engineers of Japan,
vol.61 No.3 (2007)

Red : Load working between outer ring track and balls
Blue : Load working between inner ring track and balls
Green : Load working between balls and cage pockets

Fig. 8 Dynamic mechanism analysis

Photos of authors
TUJ (54 degrees design)
Fracture torque

Conventional 50 degrees design

46

48

50
52
54
Operating angle (degrees)

56

Fig. 9 Comparison of dynamic torsion strength-19-

Manabu HOSHINO

Masashi FUNAHASHI

C.V. Joint Engineering Dept.
Automotive Sales
Headquarters

C.V. Joint Engineering Dept.
Automotive Sales
Headquarters

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Measurement of the Internal Forces ofBall Fixed Constant Velocity Joints

Daiji OKAMOTO*
Hirokazu OOBA*
In recent years, the precise measurement of internal forces inconstant velocity joints (CVJ) has become important in order tomeet requirements for lightweight, reduced size, and high efficiency.
The internal forces of a plunging-type CVJ have already beenclarified both theoretically and experimentally. On the other hand,for fixed-type ball CVJs, the internal forces had been analyzed by
employing a dynamic numerical simulation, but experimental
verification had not yet been achieved. In this study, a new test rig
was developed in order to measure internal forces in a fixed-type ball CVJ. The measured results arein good agreement with those obtained from dynamic numerical simulations. In addition, we
demonstrated the influence of the coefficient of friction, and the clearance between the cage and thesurfaces of the inner and outer rings had on the internal forces.

cage.
We also investigated the effects of the coefficient offriction of the lubricant, and the clearance acrossspherical surfaces onto each of these forces 6)-7).
Furthermore, we have compared the simulation resultsfrom the already developed general-purpose
mechanism analysis software ADAMS 4)-5) with theresults of our experimental verification.

1. Introduction
Constant velocity joints can be categorized into twogroups, that is, fixed type and plunging type. NTN has
so far established a method for simulating the force
acting inside a plunging constant velocity joint throughdynamic numerical simulation and experimental
verification 1)-3). For fixed constant velocity joints, NTN
has also developed a dynamic numerical simulation
technique for internally acting forces 2)-5). However, thespherical forms of the areas being measured including
tracks make a measuring operation extremely difficult,and as a result, experimental verification for theinternally acting forces has not yet been attempted.
To address this situation, we have fabricated a
specially instrumented outer ring that permits, bymeans of piezoelectric devices, the forces acting
inside a Rzeppa constant velocity joint (BJ) which is a
fixed constant velocity joint, to be measured. Using
this outer ring, we analyzed the forces acting on theloading side and non-loading side tracks and the balls.
We were also able to analyze the forces acting on thecontact areas between the inner spherical surfaces on
the outer ring and the outer spherical surfaces on the

2. Conditions and parameters formeasurement
The measurement conditions are summarized inTable 1, and the internal forces being measured(forces acting between the balls and tracks and aforce acting between the cage and outer ring sphericalsurfaces) are illustrated in Fig. 1, while detailed data
for the specimens used are given in Table 2. To
determine the effect of a friction coefficient, we used
two lubricants, each having a unique friction
coefficient. To investigate the effect of clearance
across spherical surfaces, we employed two cages,
each having unique internal and external spherical
diameters.

*New Product Development Dept. New Product Development R&D Center-20-

Measurement of the Internal Forces of Ball Fixed Constant Velocity Joints

Table 1 Measurement conditions
Test rig

Power circulating joint tester

Drive system

Driving inner ring (torque is transmitted
from inner ring to balls and outer ring)

Running speed (min-1)

10

Load torque (Nm)100

Working angle (deg)10，20，30，40

Test joint

BJ95L

Force acting between
balls and track
Non-loading side track load P4

terminals, the tracks and spherical surfaces wereprecision finished to ensure good dimensional
accuracy for this outer ring. In accordance with themeasured outer ring dimensions, an appropriate innerring, cage, and balls were then incorporated.
A cross-sectional view showing the installed
piezoelectric devices is included in Fig. 3. Because amoment occurs on each load-measuring terminal, due
to the shifting of contact areas between the balls andthe cage, two piezoelectric devices (see Fig. 4) werearranged in the axial direction, and a load on eachpiezoelectric device was measured.

FForce acting between
balls and track
Loading side track load P3

Outer
ring
Cage
Inner
ring

Torque direction

Ball

Rotation
direction(a) Instrumented outer ring

Fig. 2 Overview photograph of instrumented outer ring

Force acting between cage and outer ring spherical surfaces
Force acting on spherical surface F

Fig. 1 Schematic illustration of internally acting forces
Bolt
A

Table 2 Details of lubricants and cages used
Lubricant(b) Load measuringterminal of load-side track

Lubricant A (μ＝0.07 NOTE) )Lubricant B (μ＝0.10 NOTE) )Terminal support
block
Load
measuringterminal

Piezoelectric
deviceCage A
Cage

Ratio of inner spherical surface clearance to ball diameter : 0.0078
Ratio of outer spherical surface clearance to ball diameter : 0.0056

O

Cage B
Ratio of inner spherical surface clearance to ball diameter : 0.0157
Ratio of outer spherical surface clearance to ball diameter : 0.0111

NOTE) Measurement values from SRV test

A(b) Axial cross(a) Circumferential
sectional view (A-O-A)
cross-sectional viewFig. 3 Arrangement of piezoelectric devices

3. Measuring method
We have fabricated an instrumented outer ring thatincorporates measuring terminals that allow us to beable to measure the forces acting between the balls
and tracks on a fixed CVJ, and the forces acting
between the cage and outer ring spherical surfaces ona fixed CVJ. The measuring terminals each compriseof separate sections in the outer ring track andspherical surfaces as well as tri-axial force sensors(piezoelectric devices).

3.1 Instrumented outer ring
A photograph of the fabricated BJ outer ring is given
in Fig. 2. After incorporating the load measuring

Fig. 4 Photograph of piezoelectric device-21-

NTN TECHNICAL REVIEW No.75（2007）

3.2 Measurement coordinates system and symbols

angle φ= 0 deg, the balls are pushed in the furthestside in the outer ring. At the rotary phase angle φ=
180 deg, the balls are situated nearest to the outerring inlet side. By referring to this coordinates system,
we measured the loading side track loads, nonloading side track loads, and forces acting on thespherical surfaces, and compiled the resultant data.

The measurement coordinate system used isillustrated in Fig. 5. The positive direction of a load in
Fig. 5 is an arrow direction. The piezoelectric devices
are arranged in accordance with this measurement
coordinates system. The mounting angle for the
piezoelectric devices to the tracks was allowed tocoincide with the contact angle. At the same time, themounting angle for the piezoelectric devices to thespherical surface was allowed to coincide with thenormal direction. The resultant forces from theinternally acting forces were calculated based on thetriaxial forces being output from these piezoelectric
devices (see Table 3).

Far side

Inlet side
Ball position at
rotary phaseangle of 0 deg

3.3 Working angle and rotary phase angle
Working angle and rotary phase angle are definedin Fig. 6. A working angle is an inclination angle that
occurs between the outer ring axis and the inner ringaxis. The shift of the ball situated in the plane that
includes the outer ring axis and inner ring axis is thegreatest (situation in Fig. 6 (a)). At the rotary phase

Working angle

Ball position at
rotary phaseangle of 180 deg(a) Working angle

Loading side track load P3
A P3z1, P3z2
P3y1, P3y2

P3x1
P4x1
Direction of torque Fx1

P3x2
P4x2
Fx2

φ=90degand rotationO

Direction ofrotation

φ=180deg

Fy1, Fy2

φ=0degFz1, Fz2
P4y1, P4y2

A

P4z1, P4z2

Force acting
on sphericalsurface F

Non-loading side track load P4
φ=270deg(a) Circumferential cross-sectional(b) Axial crossview
sectional view (A-O-A)(b) Rotary phase angle
Fig. 6 Working angle and rotary phase angle

Fig. 5 Measurement coordinate system

Table 3 Classification and details of internally acting forces
Internally acting forces

SymbolCalculation formulas for resultant forces

Loading side track load P3
Contact force between balls andouter ring loading side track

P3x（P3x1＋P3x2）
P3y（P3y1＋P3y2）
P3z（P3z1＋P3z2）

P3＝

P3x2＋P3y2＋P3z2

Non-loading side track load P4
Contact force between balls andouter ring non-loading side track

P4x（P4x1＋P4x2）
P4y（P4y1＋P4y2）
P4z（P4z1＋P4z2）

P4＝

P4x2＋P4y2＋P4z2

Force acting on spherical surface F
Contact force between cage andouter ring spherical surface

Fx（Fx1＋Fx2）
Fy（Fy1＋Fy2）
Fz（Fz1＋Fz2）

F＝ Fx2＋Fy2＋Fz2-22-

Measurement of the Internal Forces of Ball Fixed Constant Velocity Joints

3.4 Measuring system

4. Analysis model

The measuring system that we used is illustrated inFig. 7. An output from a piezoelectric device isamplified by an inline charge amplifier, transmitted via
a slip-ring situated at the shaft end of the test rig, andis recorded by a memory high coder. After completionof the measuring operation, the obtained data isprocessed with a personal computer. With thismeasuring system, the output signal from the rotary
encoder is input every 0.1 degree (3600 pulses per
revolution), and acts as a trigger for the system toread the data. In our measuring operation, a BJ wasused as a joint that constituted a pair together with thetest outer ring. Usually, a plunging constant velocity
joint having axial freedom constitutes a pair together
with a fixed constant velocity joint such as the test
outer ring. When compared with a fixed constantvelocity joint, a plunging constant velocity joint does
not have as large of a working angle. Therefore, for
our measuring operation, the BJ was used so as to beable to obtain a working angle up to 40 degrees. Our
measuring system is constructed such that the test rig
to which the BJ is mounted, slides (white arrow in Fig.
7) to provide a working angle.

The dimensions of the test outer ring and samples
used for the experiment were measured using a 3D
measurement system. Based on the obtained shapedimensions and friction coefficients of the lubricants,
simulation was performed by the already developed
general-purpose mechanism analysis ADAMS 4)-5).

5. Results of measurement and analysis
5.1 Effect of friction coefficient of lubricant
5.1.1 Loading side track load P3
The influence of different lubricant types anddifferent friction coefficients resulting from different
lubricant types on the loading side track load P3 isillustrated in the measurement result and analysis
result in Fig. 8. The resultant findings are described
below:[1] Peak load was observed at around a rotary
phase angle at which the balls are situated in
the far side and inlet side of the outer ring.
[2] With a greater friction coefficient, the peak ofloading side track load shifts to a greater rotary

Sliding direction on test rig
Test outer ring
BJ

1 Piezoelectric
device2 Inline charge
amplifierRotary encoder
3600 P/R

4Slip-ringStart pulse 1P/R

3Output reader

5 Memory high coder(A/D conversion)

Fig. 7 Measurement system-23-

6 Arithmetic operationand indication on PC

NTN TECHNICAL REVIEW No.75（2007）

phase angle under a low working angle
situation.
[3] According to the analysis results too, thefindings [1] and [2] remain true.

below:[1] Peak load was observed at around a rotary
phase angle at which the loading side trackload decreases.
[2] With a greater friction coefficient, the peak of
non-loading side track load shifts to a greater
rotary phase angle under a low working angle
situation.
[3] According to the analysis results too, thefindings [1] and [2] remain true.

5.1.2 Non-loading side track load P4
The influence of different lubricant types anddifferent friction coefficients resulting from different
lubricant types on the non-loading side track load P4 isillustrated in the measurement result and analysis
result in Fig. 9. The resultant findings are described

Friction coefficientof lubricant

Analysis result
Force acting between balls
and track on loading side kN

Force acting between balls
and track on loading side kN

μ＝0.07

Measurement result
3
10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

3

2

1

0
0

3
10deg
20deg
30deg
40deg

1

0
0

120

240

360

Rotary phase angle deg
Force acting between balls
and track on loading side kN

μ＝0.10Force acting between balls
and track on loading side kN

Rotary phase angle deg

2

10deg
20deg
30deg
40deg

120
240
Rotary phase angle deg

360

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 8 The influence of coefficient of friction of lubricants on loading side track load P3 (Lubricant: A and B, Cage: A)
Friction coefficientof lubricant

Analysis result
10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

Force acting between balls andtrack on non-loading side kN

Force acting between balls andtrack on non-loading side kN

μ＝0.07

Measurement result
3

3

2

1

0
0

3
10deg
20deg
30deg
40deg

1

0
0

120

240

360

Rotary phase angle deg

120
240
Rotary phase angle deg

360

Force acting between balls andtrack on non-loading side kN

μ＝0.10Force acting between balls andtrack on non-loading side kN

Rotary phase angle deg

2

10deg
20deg
30deg
40deg

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 9 The influence of coefficient of friction of lubricants on non-loading side track load P4 (Lubricant: A and B, Cage: A)-24-

Measurement of the Internal Forces of Ball Fixed Constant Velocity Joints

revolution.
[2] When the friction coefficient is 0.07, the load at
the working angle of 30 deg is virtually samewith that at the working angle of 40 deg. In
contrast, when the friction coefficient is 0.10,
the peak loads are greater with the working
angle of 40 deg.

5.1.3 Resultant force from spherical surfaces
The measured force (F) which acted on sphericalsurfaces is a load acting on one of the six sphericalsurfaces on the outer ring, and the resultant force
mentioned here is a load acting from the entire cage
onto the spherical surfaces on the outer ring. By
determining the resultant from the spherical surfaces,
it is possible to determine the load acting on the cage
within a joint. The resultant force from the sphericalsurfaces can be calculated from the result of the force
acting on the six spherical surfaces. The measurement
results of the force acting on the spherical surfaces
are graphically plotted in Fig. 10. A formula for
calculating a resultant force from spherical surfaces isgiven in the expression (1) below:
Resultant force from spherical surfaces(αdeg)= (F (αdeg)2+ F (α+60deg)2+ F (α+120deg)2+ F (α+180deg)2+ F (α+240deg)2+ F (α+300deg)2)1/2……………………………………（1）where, α: given rotary phase angle
From this measured force (F) acting on sphericalsurfaces, another measurement result (resultant force
from spherical surfaces) was obtained with a different
lubricant, and another analysis result (resultant force
from spherical surfaces) was reached with anotherfriction coefficient. These new results are graphicallyplotted in Fig. 11. The so-obtained results are
summarized below:[1] From the measurement result and analysis
result, it is apparent that there are six peaks per

Friction coefficientof lubricant

Frictioncoefficientof lubricant
Force acting between outer ringand cage (sensor output) kN

μ＝0.07Force acting between outer ringand cage (sensor output) kN

μ＝0.10Force acting between
outer ring and cage kN

Force acting between
outer ring and cage kN

1

120

240

0
0

120

240

360

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

1

120

240

360

Rotary phase angle deg

3
10deg
20deg
30deg
40deg

2

10deg
20deg
30deg
40deg

2

0
0

360

Force acting between
outer ring and cage kN

Force acting between
outer ring and cage kN

1

3
10deg
20deg
30deg
40deg

2

1

0
0

2

Analysis result

Rotary phase angle deg

μ＝0.10

10deg
20deg
30deg
40deg

Fig. 10 Measurement results of forces F acting between
outer spherical surface of cage and inner sphericalsurface of outer ring (Lubricant: A and B, Cage: A)

Measurement result003Rotary phase angle deg

3

μ＝0.07

Measurement result

120
240
Rotary phase angle deg

360

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 11 The influence of coefficient of friction of lubricants on resultant forces acting between outer spherical surface of
cage and inner spherical surface of outer ring (Lubricant: A and B, Cage: A)-25-

NTN TECHNICAL REVIEW No.75（2007）[2] With a greater dimension with clearance between
spherical surfaces, and when the working angle
is smaller, the peak position (previously ataround the rotary phase angle of 0 deg) shifts toa greater rotary phase angle region.
5.2.2 Non-load side track load P4
Measurement results and analysis results wereobtained for multiple cages, each having a uniquely
dimensioned clearance between spherical surfaces.
The non-loading side track loads P4 in these results
are provided in Fig. 13. The so-obtained results are

5.2 Influence of clearance between sphericalsurfaces
5.2.1 Loading side track load P3
Measurement results and analysis results wereobtained with a plurality of cages each having auniquely dimensioned clearance between sphericalsurfaces. The loading side track loads P3 in theseresults are provided in Fig. 12. The so-obtained
results are summarized below:[1] With a greater dimension with clearance between
spherical surfaces, the peak loads are greater.

Analysis result
Force acting between balls
and track on loading side kN

Clearance between
spherical surfaces
Small

Measurement result
Force acting between balls
and track on loading side kN

Cage
3

10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

3

2

1

0
0

3
10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

Rotary phase angle deg

120
240
Rotary phase angle deg

360

Force acting between balls
and track on loading side kN

Force acting between balls
and track on loading side kN

Rotary phase angle deg

Clearance between
spherical surfaces
Large

10deg
20deg
30deg
40deg

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 12 The influence of clearance between spherical surfaces of cage and outer ring on loading side track load P3(Lubricant: A, Cage: A and B)
Analysis result

3
10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

Force acting between balls andtrack on non-loading side kN

Clearance between
spherical surfaces
Small

Measurement result
Force acting between balls andtrack on non-loading side kN

Cage

3

2

1

0
0

3
10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

Rotary phase angle deg

120
240
Rotary phase angle deg

360

Force acting between balls andtrack on non-loading side kN

Force acting between balls andtrack on non-loading side kN

Rotary phase angle deg

Clearance between
spherical surfaces
Large

10deg
20deg
30deg
40deg

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 13 The influence of clearance between spherical surfaces of cage and outer ring on non-loading side track P4(Lubricant: A, Cage: A and B)-26-

Measurement of the Internal Forces of Ball Fixed Constant Velocity Joints

summarized below.
Let us consider a situation with a greater clearance
between the spherical surfaces. At a working angle of
40 deg and a rotary phase angle of 120 deg, a nonloading side track load is detected. The analysis result
fairly well matches this finding.

of friction of the lubricant used, and the dimension of
the clearance between spherical surfaces. As a resultof our verification work, the optimized design for a
fixed constant velocity joint can be now analyzed
easily and accurately in a shorter time span.

5.2.3 Resultant force from spherical surfaces
A procedure similar to that described in Sec. 5.1.3
was performed. The resultant forces acting between
the spherical surfaces (F) as well as the influence of
resultant forces from the spherical surfaces aregraphically plotted in Figs. 14 and 15. The soobtained findings are summarized below:
With a greater clearance between the sphericalsurfaces, the resultant force from the sphericalsurfaces is lower and the vibration amplitude
increases. The analysis result fairly well matches thisfinding.

Cage
Force acting between outer ringand cage (sensor output) kN

Clearance between
spherical surfaces
Small

Force acting between outer ringand cage (sensor output) kN

Clearance between
spherical surfaces
Large

We have not only developed a unique measuring
technique for determining forces acting inside fixed
constant velocity joints (previously extremely difficulttask) but also verified reliability of analysis result thatare obtained from the already developed dynamic
general-purpose mechanism analysis softwareADAMS. In addition, we have clarified that the forces
acting within any fixed constant velocity joint are
affected and varied by the magnitude of the coefficient

Measurement result
Force acting between
outer ring and cage kN

Force acting between
outer ring and cage kN

1

120

240

360

120

240

360

3
10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

10deg
20deg
30deg
40deg

2

1

0
0

120

240

360

3
10deg
20deg
30deg
40deg

2

Force acting between
outer ring and cage kN

Force acting between
outer ring and cage kN00

Rotary phase angle deg

3

1

0
0

1

3
10deg
20deg
30deg
40deg

Rotary phase angle deg

Clearance between
spherical surfaces
Large

2

Analysis result

2

0
0

10deg
20deg
30deg
40deg

Fig. 14 Measurement results of forces F acting between
outer spherical surface of cage and inner sphericalsurface of outer ring (Lubricant: A, Cage: A and B)

3

Clearance between
spherical surfaces
Small

3

Rotary phase angle deg

6. Conclusion

Cage

Measurement result

120
240
Rotary phase angle deg

360

10deg
20deg
30deg
40deg

2

1

0
0

120
240
Rotary phase angle deg

360

Fig. 15 The influence of clearance between spherical surfaces of cage and outer ring on resultant forces acting between
outer spherical surface of cage and inner spherical surface of outer ring (Lubricant: A, Cage A and B)-27-

NTN TECHNICAL REVIEW No.75（2007）

References
1) Kei Kimata: Friction and lubrication in constant velocity
joints, Lubrication, vol.31 No.10 pp.697-702 (1986)
2) Kei Kimata, Haruo Nagatani, Masayuki Imoto, Takeshi
Kohara: Numerical Analysis and Experiments on the
Characteristics of Ball-Type Constant-Velocity Joints,
Transactions of the Japan Society of MechanicalEngineers. C, vol.69, No.678 pp.487-495 (2003)
3) Nozaki, Kohara: Experimental determination of forces
acting on tripod constant velocity joint housing, theJapan Society of Mechanical Engineers, 2003 annual
meeting preceedings, pp.155-156 (2003)
4) Yoshihiko Hayama: Dynamic Analysis of Forces
Generated on Internal Parts for Ball-Type Constant
Velocity Universal Joint; Double-Offset-Type and BallFixed-Type, JSAE Transactions vol.34, No.4 pp.154162 (2003)
5) Yoshihiko Hayama: Dynamic Analysis of Forces
Generated on Internal Components of a Double Offset
Constant Velocity Universal Joint (DOJ) by Utilizing
ADAMS software, NTN Technical Review, No.69,
pp.97-102 (2001)
6) Okamoto, Hayama: Measurement of internal forces in
fixed constant velocity joints (effects on friction), Japan
Society of Mechanical Engineers Machine Design andTribology Division Lecture Meeting, Meeting
Proceeding, pp.151-152 (2006)
7) Okamoto, Akamatsu: Measurement of internal forcesin fixed constant velocity joints (effects on friction),
Tribology Conference Preprints, pp.139-140 (2007-5)

Photos of authors

Daiji OKAMOTO

Hirokazu OOBANew Product Development Dept.
New Product Development
R&D Center

New Product Development Dept.
New Product Development
R&D Center-28-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Technical Trends in Axle Bearings andthe Development of Related Products

Kiyotake SHIBATA*
Takayuki NORIMATSU*

Demand for lighter weight, more compact, and lower friction
automotive parts is increasing rapidly with the purpose of
improving fuel efficiency. NTN has developed various products
in order to meet these demands. This report introduces these
efforts for axle bearings.

1. Introduction

2. Topology technique-based superlight hub bearing

Among environmental issues, global warming
seems to be responsible for the increased ambient air
and seawater temperatures around the earth. Natural
disasters seemingly resulting from global warming are
occurring around the world. The Kyoto Protocol, which
was passed in December 1997 and enacted in 2005,
requires each member nation to reduce greenhousegas emissions.
A major greenhouse gas is carbon dioxide (CO2).
To help reduce CO2 emissions the automobile
industry around the world has increased attempts toimprove the fuel economy of their cars by employing
lighter and lower-friction designs for their automotive
components.
NTN has also been committed to this effort byoffering lighter, lower-friction automotive components.
This paper will discuss the following four points
associated with the above-mentioned efforts:

Through the optimization of the hub flange shape
we have minimized the weight of hub bearing design,
thus leading to the successful development of a“super light-weight GEN3 hub bearing for economy
vehicles” whose mass is only 1 kg (a 30% reduction
compared to a conventional design).
We have applied our topology optimization
technique to light-weight designs for GEN3 hubbearing products in an attempt to establish atechnique that allows any person to obtain the optimal
shape and succeeded in developing a significantly
lighter hub bearing design.

2.1 Overview of topology optimization
Using the basic approach of topology optimizationmaterial is added to a key area while the thickness is
reduced from less functional areas in order to attain alean shape, while still maintaining mechanical strengthand rigidity requirements.
The shape created using the topology optimization
technique is usually a very complicated 3D shape.
Consequently, a product with this shape cannot easily
be produced with a standard industrial process and as
a result it may be unjustifiably expensive. Therefore,
we have to use a result of topology optimization as astarting point to develop a final and easy-to-machine
shape.

(1) Topology technique-based super-light hub
bearing(2) Positively sealed & low friction hub bearing w/
built-in seal
(3) High load capacity tapered roller hub bearing(4) GEN4 (4th generation) hub joint*Axle Unit Engineering Dept. Automotive Sales Headquarters-29-

NTN TECHNICAL REVIEW No.75（2007）

Through topology optimization, we are able toanswer the question “what is an optimal structure for agiven part so that the part can achieve its necessaryfunctions?” This technique can assist in changing thestructure of a mechanism dramatically.

executing an analysis operation (Fig. 1). The objective
and constraint conditions applied are given below:
Objective: Strain energy minimization
Constraint conditions: 22.5%, 27.5% and 30%
volumes relative to the volume, or mass, of the initial
shape(1) Analysis model(2) Analysis result

The analysis model used is illustrated in Fig. 1.
Assuming that the hub spindle and inner ring are asingle piece, a mesh topology was developed on thisblock to restrict the translation 3 degrees of freedom
for the nodes at the bolt holes on the flange. As for theexternal load acting on the hub spindle, the loadconcentration was assumed to occur on the rollingsurface by employing a rolling element load value thatis obtained using a bearing internal force analysis
program. At the same time we employed a rather
larger region for a possible shape change (design
region) and defined a region where shape change is
impossible due to the mounting location on acounterpart member (non-design region), thereby

The shapes in Fig. 2 and the maximum stress data
in Fig. 3 were obtained through the analysis operation.
Fig. 3 shows how the stress greatly increases with anattempt for a lighter design. We have selected theshape at 27.5% (relative to the initial volume), which is
closest to the threshold stress value as the basis forbasic structure of the hub bearing.

2.2 Topology optimization

Intermediate dia.section
Outer side: non-design
regionInner side: design region

Non-design region forbolt-fastened area

Shaft: non-design region

Inner ring
Wheel spigot
Outer circumference:
non-design region

Appearance viewed
Appearance viewed
from inner sidefrom outer side
Fig. 1 Topology optimization modelThreshold stress value

Design region
for flange

We also applied the above-mentioned technique forshape optimization of the outer ring. At the same time
we achieved a light-weight arrangement to the bearing
internal design by the development of materials andnew grease, thereby finalizing the bearing

Max. stress MPa

Hub spindle

2.3 New design

10

20

40

Fig. 3 Principal stress of each design

Shows the initial form
(See Fig. 1.)

Outer circumference of wheel spigot(Non-design region)

30

Ratio of initial shape to volume ％30% of the initial volume

27.5% of the initial volume

22.5% of the initial volume

Fig. 2 Result of topology optimization-30-

Technical Trends in Axle Bearings and the Development of Related Products

specification. The hub bearing we developed has theshape shown in Fig. 4. The final design attains thetargeted mass of 1.0 kg while maintaining the desired
strength, durability, and rigidity. The hub spindle andouter ring achieved weight reductions that aresummarized in Fig. 5.

3. Positively sealed & low friction hub
bearing w/ built-in seal
Approximately half of rotational friction (rotational
resistance) on a hub bearing is accounted for bysliding resistance of the seal. Reduction in sliding
resistance of the seal contributes to greatly improved
fuel economy of a given vehicle.
However, reduction in sliding resistance of the seal
can also lead to poorer sealing performance. This
creates a challenge for reducing rotational friction andimproving seal performance.
NTN has created a seal that performs two times
better than a conventional seal in regards to muddy
water resistance and has a 15% reduction in rotational
friction (30% reduction with seal alone) while
maintaining the conventional seal space by: [1]adoption of low friction rubber material, [2] low friction
design through optimization of the radial lip, and [3]improvement in muddy water resistance byincorporation of double side lips design.

3.1 Features

Fig. 4 Shape of final design

Conventional design

With a cross-sectional area same as that of theconventional seal:[1] Sealing performance (muddy water resistance):
two times as great[2] Rotational friction of hub bearing: 15%
reduction (30% with seal alone)

New design
Outer ring

3.2 Structure−100g

We adopted a low friction material with a uniquerubber blend for the seals installed on the hub bearing
in order to reduce rotational friction caused by the seal
lips. We have also optimized the radial (grease) lip
and introduced double side lips to satisfy the
conflicting needs of improved sealing performance
while reducing rotational friction (Fig. 6).

Hub ring−240g

Fig. 5 Amount of each component lightening

Conventional
designNew
design[Improved sealing]
Double side lips
arrangement

Side lip
Radial lip[Low friction design]
Optimized radial lip
Adoption of low friction rubber material

Fig. 6 High-sealing, low-torque hub bearing structure-31-

NTN TECHNICAL REVIEW No.75（2007）

3.3 Performance

4. High load capacity tapered roller
hub bearing

In Figs. 7 and 8 the conventional design iscompared with the new design by examining theresults of muddy water resistance test and rotational
friction test.
From the data in Figs. 7 and 8 we can see that whencomparing the conventional design with the new
design, the muddy water resistance is improved twotimes and there is a 15% reduction in rotational friction
within the entire bearing (30% reduction with sealalone).

Passenger cars use hub bearings whose rolling
elements are steel balls, while many pickup trucks and
SUV’s use tapered roller hub bearings whose rolling
elements are tapered rollers. Tapered roller bearings
have a greater load capacity and higher rigidity whencompared to ball bearings.
NTN has thoroughly re-evaluated the previous
internal design specification for it’s tapered roller hub
bearings and has successfully developed a uniquehigh load capacity tapered roller hub bearing. The
number of the rollers can be maximized(approximately same as with a full complement roller
bearing) without reducing the cage strength bysituating the cage nearer to the outer ring in order toreduce the roller-to-roller clearances; this results in thesize of the new tapered roller hub bearing being thesame as the old design. This new design has helped
increase the bearing life by 20% and the bearing
rigidity by 7% without increasing the bearing size.

Muddy water resistance cycles

Muddy water resistancetest results

Two times as great

4.1 Features
Conventional
design[1] Longer bearing life: 20% increase in life
(calculated life)[2] Greater bearing rigidity: 7% increase in bearing
rigidity (comparison to old design)New
designFig. 7 Muddy water resistance test results
Rotational friction
test results

4.2 Structure
15% reduction
in rotational
friction

The cage diameter is increased towards the outercircumference in order to realize a smaller roller-toroller clearance dimension. The result of this is thatthe number of rollers on the new design can be as
great as that of a full complement roller bearing of asimilar size (Figs. 9 and 10).

Rotary friction

Seal
sectionBearing
sectionConventional design
Conventional design

Clearance between
outer ring and cage

New design

Width of pillar
on cageFig. 8 Rotational friction test results

Roller-to-roller
clearance

High load capacity tapered roller hub bearing
Clearance between outerring and cage (smaller)

Width of pillaron cage
(Same as that withthe conventional
design)Roller-to-roller
clearance (smaller)

Fig. 9 High-load capacity tapered roller hub bearing structure-32-Technical Trends in Axle Bearings and the Development of Related Products

outside diameter on the outer ring pilot section of the
new design is 7 mm (6%) smaller and the resultant
weight reduction is 400 g (6%).

Conventional design

5. GEN4 hub joint

Wider

NTN has developed a 4th generation hub joint
product series (GEN4 H/J) by integrating a constantvelocity joint (CVJ) with a hub bearing (H/B). This
product line contributes to lighter weights, morecompact sizes, higher rigidity and easy assembly
procedure for axles. The structural features and basic
characteristics of the GEN4 H/J products aredescribed below.

High load capacity tapered roller hub bearing

5.1 Structure
NTN’s newly developed tube forming method isapplied to GEN4 H/J as a method for fastening theH/B to the CVJ outer ring. This tube forming method
for the GEN4 H/J is described below:[1] By knurling, square pyramid shaped knurls areformed inside the bore surface of the hub.
[2] The entire area containing the knurls is hardened
through heat-treatment.
The appearance of the knurled area including theknurls is shown in Fig. 12.
[3] A punch is driven into the bore of the stem on the
CVJ outer ring to expand the stem in thecircumferential direction, so that the stem is formed
to the knurls (Fig. 13).
The area where the stem is formed to the knurls ishereinafter referred to as the expansion-formed
section.

Narrower

Fig. 10 Cage part appearance

4.3 Design study example
As described above, the new high load capacity
tapered roller hub bearing will offer a longer life and ahigher rigidity compared to that of a conventional
design. The new bearing design can be more compact
and light-weight without the need for an increase in
the bearing diameter.
One design study example is shown in Fig. 11. The

Specification

SchemeLightnessBasic
diameterBasicweight

−7mm

−400g

（6%）

（6%）

φD

Conventional
designCompactness
（φD）

φD

New
designFig.11 High-load capacity tapered roller hub bearing study-33-

NTN TECHNICAL REVIEW No.75（2007）

Punch

Fig. 12 Knurl appearance

Fig. 13 Tube forming method

5.2 Advantages of new design
Another feature of the GEN4 H/J is that the raceway
surface of the H/B is directly formed on the CVJ outerring. At the same time, by adopting a high efficiency
CVJ (EBJ), the GEN4 H/J offers the following
advantages:

structured so that the entire H/J assembly can to passthrough the bore of the knuckle (Fig. 14).
Consequently, component handling on the vehicle
assembly line and serviceability in the market are
improved.

(1) Light-weight, compact design
The weight is reduced which in turn contributes toimproved handling, stability of the vehicle, and better
fuel economy.

5.3 Design study example
The structure and advantages of the GEN4 H/J
have been presented so far. Concrete data forpossible reduction in weight and size is described
below, which refers to a study using a large-sized FF
layout car with an engine displacement of 3 liters (Fig.
15).
As summarized in Fig. 15, compared with aconfiguration that is based on a conventional GEN3
H/B and CVJ the new design contributes to an 800 g(14%) weight reduction per wheel and a 22 mm (20%)
size reduction per wheel.

(2) Improved degrees of freedom for suspension
designA shorter axial distance between the hub flange andthe CVJ helps to increase the degrees of freedom forsuspension design. The GEN4 hub joint can beapplied to a suspension of a complicated structure tohelp improve handling and stability of the vehicle.

(3) Easy assembly and improved serviceability
The GEN4 H/J assembly (outboard-side fixed H/J,
intermediate shaft and inboard-side plunging joint) is

Knuckle inside diameter

Fig. 14 Easy assembly and improved serviceability-34-

Technical Trends in Axle Bearings and the Development of Related Products

Specification

SchemeLightness

Compactness※BasicweightBasiclength

−800g
（14%）

−22mm
（20%）

Conventional
designGEN3 H/B＋
CVJ
L

Expansionformed typeGEN4 H/J

L*Comparison in terms of the axial distance between the hub flange and CVJ center.

Fig. 15 GEN4 H/J study

5.4 Evaluation test

References

The following bench tests were performed to verify
that the strength and durability of the expansioncaulked section are at least equivalent to those of nutfastened section and oscillation-formed section.
Bench tests(1) Pulling force measurement(2) Static torsional strength measurement(3) Torsional fatigue strength test(4) Rotary bending test
In addition, to evaluate the overall durability of theGEN4 H/J we have successfully tested this line ofbearing products on actual cars (overturning moment
test, full steering & violent start test, etc.) without any
issues.

1) Haruo Nagatani, Tsuyoshi Niwa: NTN Technical
Review, No.73 (2005)
2) Shigeaki Fukushima, Hiroyuki Ogura: NTN Technical
Review, No.70 (2002)

6. Conclusion
We have presented our efforts for reduction inweight and friction with axle bearings in order toimprove fuel economy and to help reduce pollution.
Though a lighter weight is an important requirement,
the axle bearings must maintain their basic functions
by supporting the axles and smoothly running.
To satisfy the ever mounting needs for better fueleconomy, NTN will fully utilize the items mentioned
above and commit to development a novel technologythat ensures a highly reliable and robust design
practice.

Photos of authors-35-

Kiyotake SHIBATA

Takayuki NORIMATSU

Axle Unit Engineering Dept.
Automotive Sales Headquarters

Axle Unit Engineering Dept.
Automotive Sales Headquarters

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Development of High Resolution Sensor Element MPS40S
and Dual Track Magnetic Encoder for Rotational Speedand Position Measurement
Pascal DESBIOLLES*
Achim FRIZ*

Higher precision and resolution requirements for rotational speed
and position sensors in automotive and other applications lead SNR toinvestigate in the potential of their ASB 1) ®-Technology, where
magnetic single track encoders have been integrated in wheel
bearings and serve together with small active sensors for detecting the
speed of each individual wheel of a car. Such information is used for
ASIC and Magnetic Encoder
ABS 2), ASR 3) , ESC 4) and other car systems.
Based on this well proven technology, which is installed in over 50 million cars now, SNR has developed
two new basic components to improve the performance of such sensing systems. The magnetic sensor
ASIC 5) MPS40S in combination with a new dual track magnetic encoder provides high resolution speed
and incremental position signals, with direction of rotation and additionally the possibility of reference orindex signals for absolute position determination.

1. Introduction:
The industrial developments go more and more into
electric devices and even traditional mechanicalsystems become electrically assisted, such as manual
gearboxes, steering systems etc…. So the need for
precise and reliable sensor is continuously increasing.
At the same time the cost factor becomes more andmore important. Since SNR has introduced the ASB ®
Technology for wheel speed detection in cars, whichbecame a world standard now, the research efforts inthat domain have lead SNR to develop not only
magnetic encoders with better magnetic material andhigher precision, but also sensing elements in order toimplement more functionality in the Sensor Bearings.
At the SNR Mechatronics Department in Annecy,
France, basic developments and research programs
are conducted especially for magnetic sensor devices
as these are seen as robust and cost effective
solutions, which can work even in harsh environment
situations. The performance of the latest sensor
development is presented in this paper as well as ashort overview of the technology.

ASIC MPS40S
Magnetic Encoder

Fig. 1 Section of high resolution sensor bearing

1) Representing "Active Sensor Bearing" named by SNR.
2) "Anti-lock Brake System".
3) "Anti Slip Regulator" or "Anti Skating Regulation system".
4) "Electric Stability Control"
5) "Application Specific Integrated Circuit".

*SNR Mechatronics Research & Development-36-

Development of High Resolution Sensor Element MPS40S and Dual Track Magnetic Encoder for Rotational Speed and Position Measurement

it is shown that such phase shift changes, resulting ina reference pulse signal, can be put on every single
pole pair of a north and a south pole. The poles are
separated by a straight transition line, when a signal
shift from 1 to 0 or from 0 to 1 is implemented. On thissample a special pattern of 65 pole pairs is
implemented. The position of the reference pulses has
been calculated according to the application
specification. Theoretically it is possible to put asmuch reference pulses as pole pairs on a magnetic
encoder of this type. SNR developed a special
magnetisation process to improve the pitch error thatis very important for such a device. The phase shift of
the transitions over one pole pair in Y/Z direction isdepending on the calculated nominal airgap (Xdirection) in order to get acceptable mounting
tolerances. Field simulations helped to optimise themagnetisation tool, which is specific for different
reading diameter and airgap. The standard
magnetisation tools do not provide sufficient precision
for the dual track encoder.

2. Sensor Components:
2.1. Magnetic Encoder
At the end of the eighties, SNR developed a firstmagnetic multi-pole encoder, which was integrated ina sensor bearing. Meanwhile such encoders are to befound in many different sensor bearings typically in
combination with standard Hall or MR elements. Thiswas one of the basic components of the ASB ®
system. The qualification program for the magnetic
encoders especially for automotive wheel applications
proved the robustness of the design; none of water,
salt, mud or small metallic particles could disturb thequality of the signal. Today SNR is introducing a newgeneration of magnetic encoders with a second trackin the same space of the single track encoders. With
the second track it is possible to put one or morereference points on the encoder, which indicate anangular position of the encoder. The magnetisation
process has been adapted for such dual track
encoders, as the magnetic field has to be wellcontrolled in order to minimise the influence from themagnetic poles of the reference track to the highresolution track. Fig 2 shows the different types ofmagnetic encoders and the corresponding magnetic
fields. Following the points 1 to 6 show the variationof the magnetic field strength, which is seen by thesensor. Standard Hall or MR (magneto resistor)
elements cannot interpret such field configurations
and would lead to a false signal. Therefore SNR has
developed a new sensing element which shall be
described paragraph 2.2.
In Fig 3 is shown how the magnetic poles look likeon a real encoder. At the inner circle of the magnetic
encoder the poles are all the same size. On the outercircle this is not the case; the red points indicate thepositions where the reference pulse is generated. The
poles are visualised with a special magnetic film, sothat the two tracks with the phase shift change
between the poles can been seen easily. Furthermore

2.2. Magnetic Sensing Element MPS40S
The magnetic field sensor ASIC MPS40S integrates
a first array of Hall elements for high resolution
purposes and a second array which reads the second
track of above described magnetic encoders in orderto create reference pulses. On the MPS40S block
diagram, Fig 4, some of the features are directly
shown. In order to make this device as flexible aspossible, some features have been implemented
which are patented. The MPS40S can read a wide
range of magnetic pole widths, which is newSingle track encoder

Dual track encoder

Fig. 3 Magnetic encoder with reference track
and signal status

Fig. 2 Scheme of single and dual track encoder-37-

NTN TECHNICAL REVIEW No.75（2007）

compared to magnetic wheel speed sensors, the polesize can be programmed which makes the readingeven more precise. A very special interpolator blockon the die of the MPS40S allows raising the resolution
over one pole pair. The interpolation factor is also
programmable as shown in Table 1. Furthermore the
MPS40S provides the direction of rotation from 90°
phase shifted signals A and B. In parallel suchmagnetic encoders, which may be integrated in abearing seal, have been designed by adding a second
magnetic track in the same envelope of the series

product. The second track is read by the MPS40S
simultaneously and provides angular position
information of the magnetic encoder. The
magnetisation process is done in a way that evenexisting designs of magnetic encoder seals can bemagnetised with the new dual track for existingproduct improvement. In this case only the sensor
head has to be adapted for integrating the MPS40S.
For a better understanding the signals are shown in
Fig 5, where the analogue signals input from bothtracks are compared to the digital signals. Reading

Fig. 4 MPS40S block diagram

Table 1 Interpolation factor programming table
Interpolation

1x

2x

4x

5x

8x

10x

16x

20x

32x

40x

Pulses 1)

32

64

128

160

256

320

512

640

1024

1280

Edges 2)

128

256

512

640

1024

1280

2048

2560

4096

5120

Resolution 3)

2.8˚

1.4˚

0.7˚

0.5˚

0.35˚

0.25˚

0.18˚

0.12˚

0.08˚

0.07˚

1) Based on a 32 pole pair encoder 2) Number of raising and falling edges of the signal output lines
3) Resolution, when using all edges of the signals

Fig. 5 MPS40S signals-38-

Development of High Resolution Sensor Element MPS40S and Dual Track Magnetic Encoder for Rotational Speed and Position Measurement¡Test bus, for application development purposes
The output of the test bus is programmable foranalogue and digital signal readings.
¡AEC Q100 qualificationfrom left to right is the clockwise direction and fromright to left the counter clockwise direction of rotation,in any case the reference pulse comes at the same
position by internal comparison of the high resolution
input and the reference input.

3. Applications and some examples:

Features of the MPS40S combined with Dual Track
Magnetic Encoder:¡Non contact sensor concept¡Compact design, smallest reading diameter:
φ35mm¡Temperature range: -40˚C to +125˚C¡Compatible with harsh environment, dust, mud,water, oil …
A lot of tests have been done, when this encoder
material has been qualified for automotive
applications such as encoder for wheel bearings incars. All car manufacturers have released our
products for such application.
¡Wide range of magnetic pole width: 1.15mm to
6mm, with a pitch of 0.02mm
Over the pole width, combined with the interpolation
factor, almost all resolutions can be achieved. The
pole width is programmable.
¡Flexible design of magnetic encoders, numberof pole pairs, …
¡Automatic Airgap Adjustment, within thedetectable magnetic field (≥5mT)
In case the detectable magnetic field becomes too
small, the MPS40S sets an alarm.
¡Programmable Interpolation Factor: 1x to 40x
The different interpolation factors are shown in
Table 1.
¡Direction of rotation information
By reading the A- and B-signal, which are 90°phase shifted. The raising edge which comes first,
whether on A or B indicates the direction of rotation(shown in Fig 5).
¡Reference pulse information (1 or more on themagnetic encoder)
The MPS40S compares the magnetic fields of bothtracks, shown in Fig 5. On the C-signal line a pulse
is sent, when at the Zero-crossing of the A-signal
the positive field of the RP-track is above thethreshold, as shown on the red line in Fig 5.
Obviously this can be repeated on every pole pairon the magnetic encoder.
¡Self diagnostic features integrated in the
MPS40SAs already mentioned, the MPS40S can detect
failures concerning the magnetic field out of thedetectable range, but also internal failures. In thiscase the MPS40S sends a failure message via thethree output lines.

Such a sensor concept can be used in manyindustrial applications, where speed and position haveto be measured on hollow shafts or where the end of arotating shaft cannot be accessed. As the sensor headis very small and the encoder can be integrated in any
rotating part, such as a bearing, applications with
reduced available space may be interesting as well.
Steering Angle Sensor, a joint development with
Continental Automotive Systems:
Sensor and encoder may be separated or
integrated, both designs are under development.
First series production developments have been
started. In Fig 6a the separated version of thesteering angle sensor is integrated lower part of thesteering system. This is the most effective design interms of space and cost and it can be even
improved if the magnetic encoder can be integrated
in a bearing, which is shown in Fig 7. Therefore thebearings must have a section, which is big enoughto get the encoder, and there must be a "window" tothis section where the sensor can read the twomagnetic tracks.
For more conservative designs the angle sensor
can be packaged in a plastic housing, where airgap
is controlled by internal design, shown in Fig 6b.
The sensor is mounted with screws and the steering
shaft goes through the sensor. This takes more
space and makes replacement more complicated,as the steering shaft has to be disassembled.
Compared to this, the separated design is moreconvenient as the electronic part of the systems canbe taken out easily and the magnetic part, which ismounted to the steering shaft, is not affected. In
each case the sensing concept is contactless anddoes not add any torque to the steering system,
neither any noise. Both points are very important forsuch applications as noise and torque are directly
felt by the driver.
EPS motor position sensor:
Sensor and encoder are separated; the encoder is
integrated in a ball bearing
The same sensor concept can also be used in therotor position sensor for electric motors andactuators, such as EPS motors. Because of the highresolution capability of the sensor, this application is-39-

NTN TECHNICAL REVIEW No.75（2007）

Sensor with
MPS40S

Dual Track
EncoderFig. 6a Steering angle sensor
SNR/Continental automotive systems

Fig. 6b Steering angle sensor

Sensor with
MPS40S

Dual Track
EncoderFig. 7 EPS motor with sensor and encoder bearing

The flexibility of the sensor with the MPS40S helps
to accelerate the development of new motors as the
programming is done through the Vdd pin. It is very
easy to test new configurations without changing
the mechanical environment.

interesting, especially if no sensor can be placed at
the end of the motor shaft. The magnetic encoder
pattern on the second track is then designed to getabsolute position over one electrical period andrepeat this pattern for all pole pairs of the motor.
Patents have been filed for the special way of thepattern layout for such applications. Even if thesensor is not absolute true power on, with thespecial pattern the absolute rotor position can be
determined within a small angle at power on andfrom this point the sensor works like an absolute
angle sensor. The enormous advantage is again
space and weight saving compared to any resolver
technology. Furthermore if the motor design has tobe changed in terms of number of pole pairs on therotor, the sensor design does not change, only themagnetic encoder gets a different magnetisationand the programming of the chip may be adapted.

High Resolution Sensor Bearing for Industrial
Applications
The demand for sensor bearings in Industrial
Applications is also increasing and the SNR
technology can be applied easily in the same way.
However the design should be more integrated toease assembly and therefore SNR created a first
sensor bearing design which uses all functions of
the ASIC and fits into ISO dimensions regarding
inner and outer diameter of a 6203 deep grove ballbearing. This bearing is available as a prototype as
the engineering department at SNR prefers to-40-

Development of High Resolution Sensor Element MPS40S and Dual Track Magnetic Encoder for Rotational Speed and Position Measurement

develop such type of bearing as an integral part of acomplete system optimised for the final applicationin terms of environmental requirements, assembly,
space and cost. Our experience shows that too
often a standard design does not perfectly fit in thedesign of the application and we are forced tomodify the "standard" product. Today SNR has all
the necessarycomponents inorder to develop
rapidly the
product which
fulfils thecustomerrequirements at
an attractive
price.
Fig. 8 Sensor bearing

4. Conclusions
The newly developed technology by SNR is mainly
designed for mass production systems, where thefocus is on reliability in first place and seconds thecost. Hall Effect based sensor technology is proven inindustry since a very long time and the improvementsthat SNR has integrated in the ASIC MPS40S makefull profit of that background. Furthermore the very
long experience in magnetic encoders for applications
such as automotive wheel bearings, where water,mud, high and low temperatures are the normal
environment, is the perfect base for SNR, whoinvented that technology and introduced it to themarket in 1997, to make the next step in magnetic
sensing technology. The very flexible combination ofboth components by programming of the MPS40S andadapting the magnetic encoders in terms of number ofpole pairs, size etc makes it easy to find a solution.
The engineering departments at SNR have all thetools to develop the integration of such sensor devices
in many applications, even close to strong magnetic
fields from electric motors or other harsh environment
conditions. With all this benefits, this technology has avery good market potential in many industrial
branches.

Radial Sensor – Incremental High Resolution
Signals
Finally the last product developed as a standard
sensor without bearing, SNR developed the Radial
Sensor, which is also using the SNR ASIC. It
delivers incremental speed and angle measurement
signals. It is not foreseen that this encoder provides
the Reference Pulse signal output; however fullprogramming capabilities concerning pole width andresolution are implemented. This sensor works
together with radial single track magnetic multi-pole
encoders, which can be found on the market. The
electrical interface is designed to work with different
supply voltages from 5 to 8-30V and the output
stages are realised as Push/Pull.

Fig. 9 Radial sensor

Photos of authors

Pascal Desbiolles

Achim Friz

SNR Mechatronics
Research & Development

SNR Mechatronics
Research & Development-41-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Tube Forming Simulation in a Generation 4 Hub joint

Akitoshi IMOU*
Takehiro TAKANO*

Accurate CAE analysis is a very useful tool for shortening
development timing and reducing the frequency ofdevelopment trials. In this report, we introduce the process ofimproving CAE accuracy in plastic forming, which is an NTN
original method of combining HUB and CVJ.

prototypes has allowed the development work to begreatly accelerated. In this paper, we present thecontent of a FEM-based analysis technique developed
for the novel tube forming method used as a fastening
work for GEN4 H/J.

1. Introduction
The need for lighter automotive components has
been ever mounting in order to realize moremaneuverable, less energy-consuming cars. Improved
functionality, lighter weight and reduced size havebeen required for wheel bearings. In this context, NTN
developed the first generation hub bearing (GEN1
H/B) that combined two previously separate bearings.
The GEN2 and GEN3 hub bearings (H/B) eachincorporated not only a stem (hub ring), which is aperipheral component of a bearing, but also a portionof housing. The NTN GEN3 H/B products are widelyused by car manufacturers today 1).
Aiming at lighter weight and more compact size,
NTN has developed the fourth generation hub joint(GEN4 H/J) that combines a constant velocity joint(CVJ) and a GEN3 H/B with each being a bearing
peripheral component 2).
H/B product design requires strength analysis by afinite element method (FEM). This practice is very
effective in decreasing the number of prototypes,
shortening the development phase, and reducing themass of the product 3)-4).
For the GEN4 H/J, we have introduced a novel
fastening technique called the “tube forming method”.
Without FEM analysis extensive prototyping would benecessary to verify reliability of this technique. FEM
analysis work prior to the fabrication of actual

2. Overview about the newly developed
tubeforming method
The newly developed technique of tube forming isutilized by NTN to fasten the hub ring to the CVJ outerring (Fig. 1). The basic process of the tube forming
method can be summarized as follows:
1) Inside the hub bore, square pyramid shaped knurls
are formed through a process known as ‘knurling’.
2) The knurls are hardened by heat-treatment.
3) A punch is driven into the stem bore on the CVJ
outer ring to expand the stem in the circumferential
direction so that the outer circumference of thestem is engaged with the knurls (Fig. 2). As aresult, the stem is fastened to the hub ring securely
and the bearing is preloaded in an appropriate
range.
The section where the knurls engage with the stemis hereinafter referred to as the "expansion-caulked
section". The appearance of knurls formed inside the
hub bore is shown in Fig. 3.

*Automotive Engineering Dept. Automotive Sales Headquarters-42-

Tube Forming Simulation in a Generation 4 Hub joint(a) Hub rings and CVJ outer rings are elasto-plastic
solid (s).
(b) The punch was treated as rigid due to its stiffness
relative to the stem.
(c) Nodes on the hub ring flange, hub ring itself, andCVJ outer ring surface were fixed in the axial
direction.
(d) Contact were defined at locations where the hub
ring is in contact with the CVJ outer ring, andwhere the CVJ outer ring is in contact with thepunch.

CVJ outer ring

Hub ring

Expansion-caulked
sectionStem section

3.2 Results from the initial model
Bearing outer ringAt first a static FEM analysis method was applied. Abalancing condition was to be determined for eachstep in engagement process with the knurls.
Consequently, an unjustifiably long calculation time
was needed. To address this problem, we attempted
to determine whether a dynamic explicit method could
be used for the analysis. The result obtained from a
dynamic explicit method on the analysis model in Fig.
4 is illustrated in Fig. 5.
The overall trend of the engaged knurls and stem
bore deformation mode in the analysis results
matched the measurement results:[1] The error in circumferential displacement on thestem bore relative to the measurement result is 5%.

Fig. 1 Tube formed Gen4 H/J
Punch

Fig. 2 Tube forming method

Axial displacement
constraint

Pilot section
Hub ring
CVJ outer ringContact points

Travel of punch

Fig. 3 Knurl appearance

Punch outside diameter

Fig. 4 Calculation model

Hub ring

3. FEM-based analysis for tube forming
3.1 Analysis model2

As shown in Fig. 4, our analysis model comprised a
hub ring, CVJ outer ring, and punch. The punch model
consisted of its outer circumference surface that is incontact with the CVJ outer ring. Due to the symmetricstructure, the expansion-caulked section was modeledaccording to the 1/2 pitch with the knurls on thecircumference . The assumptions for the testing are as
follows:

Stem section

3

Punch outside
diameterFig. 5 Tube forming simulation result

-43-

1

NTN TECHNICAL REVIEW No.75（2007）[2] The expansion at the end of the expanded portionis outward.
[3] The axial elongation over the entire length of
expanded portion, as well as the elongation at theend of expanded portion are greater in the bore
side rather than in the outer circumference side.
The trends with the earlier models were consistent
with the measurement result, and it was apparent thatanalysis for the tube forming process was possible.
However, some factors in the analysis results, such as
excessive punch load and depth of engagement withthe knurls, did not match the measurement result. In
order to correct this, and provide more accurate
results, the model was modified.

Analysis result
Punch load

Measurement result

Time

Fig. 6 Punch load

3.3 STEP 1
To correct the aspects whose analysis results did
not match the measurement results, we attempted to
improve consistency of the punch load. If the punchload (which is a macro characteristic) in the analysis
results does not coincide with the measurement
results, then factors such as the stress in the analysis
results would also not coincide with those in themeasurement results. Since analysis for the tube
forming process is elasto-plastic, the results can varydepending on factors such as the material data in theplastic regions. For this reason, we obtained
deformation resistance curve for the hub ring.
Incidentally, for the first model, the friction coefficient
occurring between the punch and stem bore wasassumed to be 0.2. However, for the tube forming
process a lubricant is applied to the punch, and theassumed coefficient was greater than the actual
coefficient. Therefore, we varied the friction coefficient
to allow the analysis results to coincide with themeasurement results (Fig. 6).

Improved engagement

Fig. 7 Detail of tube expanding part3.5 STEP 3
Next, let us compare the expansions on the outercircumference of the pilot section. The expansion
value in the analysis results was about one half of thatfound in the actual measurement results making thevalue of circumferential stress occurring on the pilot
section lower than expected. Therefore, to investigate
he influences of unknown factors on the expansion of
the outer circumference of the pilot section, weinvestigated the following:[1] Finer mesh for the stem section:
In addition to the axial mesh in STEP 2, a meshmore finely divided in the circumferential direction
on the CVJ stem was used to simulate thedeformation on the stem section.

[2] Altered shape for the projection tips:

3.4 STEP 2

The tip shape the of knurls was measured, and theactual shape was reflected in the analysis model.

In a metal forming process such as our tube forming
process, wherein projections (knurls) could be
expected to engage with a surface, the projections
may fail to engage with the surface when the mesh is
to coarse. To address this problem, the previous
coarse mesh was replaced with a finer mesh. This
makes the projections more positively and deeply
engaged with the surface (the stem side is deformedin accordance with the contour of these projections).
However, at the same time, deeper engagement byprojections can cause the elements to have too muchdeformation and the elements can be disrupted. In an
attempt to improve the engagement situation wevaried the contact parameters. This did improve thesituation although it did not fully solve the problem.

[3] Alteration to the initial clearance on the stem
as well as to the expansion allowance:
The initial clearance between the stem outside
diameter and the inside diameter of the knurledbore as well as the expansion allowance between
the punch and the stem bore were converted into
the measurement result.

[4] Alteration to the material characteristics forthe stem section:
In addition to the deformation resistance curve forthe hub ring obtained in STEP 1, data for a similar
curve for the CVJ outer ring was obtained. The
data was then translated into the characteristics ofthe material of the CVJ outer ring. As shown in-44-

Tube Forming Simulation in a Generation 4 Hub joint

Stress HighStress Low

Fig. 10 Stress distribution at the time of tube forming simulation processing

Figures 8 and 9, after implementing the above
factors, the outside diameter expansion on the pilot
section in the analysis results now coincides withthat found in the measurement results. andSimilarly, the stress value has also improved.
Our analysis technique can be now be utilized withconfidence, and greatly reduce the number of test
as well as the overall development time.

4. Conclusion
We have developed a new analysis technique forthe tube forming process that is used as a fastening
process for GEN4 hub joint products. The resultant
improved accuracy of analysis is sufficiently high for apractical application. Through this newly developed
analysis technique, it is now possible to define a
combination of optimal forming conditions while at thesame time accelerating the development of new hub
joint designs.
The remaining challenges to be addressed will
include the development of a technique for evaluating
the mechanical strength of the joint resulting from thetube forming process, in terms of the necessary
extraction power and torsional strength. We willaddress these challenges, and further improve our
CAE technique, thus allowing it to be fully utilized inour design activities.

3.6 Stress value

Expansion on outer diameterWith our analysis technique, we are now capable ofdetermining the overall stress level on the stem
section that results from the tube forming process, andwe can now determine portions where a greater stressoccurs from the tube forming process (Fig. 10).

1Adoption of finer mesh for the stem section
2Alteration to the knurl tip shape
3Alteration to the stem initial clearance andReferences

expansion allowance1) Eiji Funahashi: History of Hub Bearings and Recent
Technology, NTN Technical Review, No.70 (2002)
2) Shigeaki Fukushima, Hiroyuki Ogura: Generation 4
Hub Joint, NTN Technical Review, No.70 (2002)
3) Akitoshi Imou: computer Analysis of Hub Bearings,
NTN Technical Review, No.70 (2002)
4) Haruo Nagatani, Tsuyoshi Niwa: Application ofTechnology Optimization and Shape Optimization for
Development of Hub-Bearing Lightening, NTN
Technical Review, No.73 (2005)

4Alteration to the material characteristics forthe stem section

Measurement First
resultanalysis

1

123

1234

Circumferential direction stressFig. 8 Quantity of outer diameter expansion

Photos of authors

1Adoption of finer mesh for the stem section
2Alteration to the knurl tip shape
3Alteration to the stem initial clearance andexpansion allowance

4Alteration to the material characteristics forthe stem section

Measurement First
resultanalysis

1

123

1234

Fig. 9 Circumference direction stress-45-

Akitoshi IMOU

Takehiro TAKANOAutomotive Engineering Dept.
Automotive Sales Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters

NTN TECHNICAL REVIEW No.75（2007）[ Technical Paper ]Development of an In-Wheel Motor Axle Unit

Minoru SUZUKI**
Kayo SAKAI**

Koichi OKADA**
Yusuke MAKINO**

In order to respond to the global demand for more energy
efficient and environmentally friendly electric vehicles, NTN
has developed an in-wheel motor axle unit for upcoming
electric vehicles. This axle unit consists of a high-reduction
cycloid and a high-speed axial gap motor to achieve acompact and lightweight design. In this paper, the initial design
explanation is presented and the on-vehicle test results areshown.

1. Introduction

2. Specification and structure of the
newly developed in-wheel motor
axle unit

Recently in Japan, not only the car industry but also
the government and private organizations have been
putting much effort into addressing global warming
issues and energy concerns. There have been
breakthrough technologies in both electric motors andrechargeable batteries for fuel cell electric vehicles(FCEV) and electric vehicles (EV) that havedramatically improved the performance of thesevehicle types 1).
The drive mechanisms for FCEV and EV can be
broken down into two categories. The first category isa more conservative system where a single electricmotor replaces a traditional internal combustion
engine. The electric motor is mounted on the chassis
and the power from the motor is transmitted to thetires via the drivetrain. The second category is an inwheel motor axle system where a separate motor isinstalled in each wheel. In-wheel motored cars arecapable of offering more passenger space than
traditional engine-powered cars and have greater
vehicle stability due to independently driven wheels.
In-wheel type designs are currently being developedby various car manufacturers 2)-4). NTN also has begun
developing an in-wheel drive unit for electric vehicles 5).
In this paper, we will explain the information learned
from the bench testing of this design as well as theresults from tests on an actual car that incorporated
our prototype in-wheel motor axle units.

2. 1 Target specification
The car that incorporated our prototype in-wheel
motor axle unit was a 1500 cc class compact car. We
desired that the car used for this testing (Fig. 1) havepower-performance comparable to current gasoline
powered cars. The targeted performance values of ourin-wheel motor axle unit are summarized in Table 1.

In-wheel motor axle unit

Fig. 1 Installation of the unit**New Product Development Dept. New Product Development R&D Center**Mechatronics Research Dept. New Product Development R&D Center-46-

Development of an In-Wheel Motor Axle Unit

Table 1 Target specification of axle unit
Max. output

20kW

Max. torque490Nm

Max. speed

150km/hMassApprox. 25 kg

Reducer type

Cycloid reducing system

Reduction ratio

1/11

Motor type

Axial gap type permanent
magnet synchronous motor

Max. motor speed

15000min-1

V planetary gear reducing mechanism that is capableof a greater reducing ratio within a smaller space. As
shown in Fig. 3, the cycloid reducer mechanism
comprises an external gear with an epitrochoidal tooth
profile, multiple internal gears each having a circular
tooth profile, and internal pins situated in the smaller
diameter portions of the external gear. The reduction
ratio can be expressed through equation (1) assumingthat the internal gears are locked in place and theinternal pins cause the rotation of the external gear.
The reduction ratio is governed by the ratio of thenumber of teeth of the external gear to that of theinternal gears. The proposed design is multiple rowed
gearing, but even single-row gearing can attain ahigher reduction ratio. Use of this gearing mechanism
causes a greater number of teeth to remain in contact,
and therefore the torque transmission per unit volume
of a reducer is greater and the reducer size can be
decreased.
Generally, the transmission efficiency of a cycloid

2. 2 Structure
Compared to other systems an in-wheel motor axle
system can be disadvantageous because the four
separate in-wheel motor axle units increases theunsprung weight, which can jeopardize driving stability
and riding comfort in these vehicles. Therefore, inorder for an in-wheel motor axle system to be effective
and customer friendly, an in-wheel motor axle unit
needs to be as lightweight and compact as possible.
The electric motor accounts for the largest amount of
the total weight in an in-wheel motor axle unit.
Generally, a specified maximum torque output governs
the motor size. Because of this specification and thedesire for a lighter unit, we have introduced a reducer
that will decrease the required maximum torque for the
motor so that a lighter motor can be used. Our in-wheel
motor axle unit schematic is illustrated in Fig. 2. The
basic components of this design consist of a hub, a
reducer, and a motor. Incidentally, the knuckle for
installing the in-wheel motor axle unit to the vehicle isalso the reducer housing.

Nout
Zi−Zo
＝−
……………………………（1）Nin
Zo
where, Nout ：output shaft speed
Nin ：input shaft speed
Zo ：number of teeth on external gear
Zi ：number of teeth on internal gears
reducer mechanism is lower than that of a 2K-H
planetary gear reducer. The lesser efficiency of thecycloid reducer seems to be due to the sliding contact
between the inner and external gears and also
between the internal pins and external gear. To
address this issue, we have attempted to reduce thepower loss by incorporating rolling bearings into thesesliding contact areas5).

2. 2. 1 Reducer Design
Instead of the commonly used 2K-H planetary gear
reducer mechanism, we adopted a cycloid reducer
mechanism 6). A cycloid reducer mechanism is a K-H-

Internal gear
External gear
Counterweight

Rotor

Water
jacketStatorHub + reducer

Input shaft

Internal pin

MotorFig. 2 Schematic of axle unit

Fig. 3 Structure of reducer-47-

NTN TECHNICAL REVIEW No.75（2007）

On an ordinary cycloid reducer mechanism, twoexternal gears (which constitute a pair) are allowed torun. The phase of one external gear is opposite that of
the other external gear in order to compensate for thevibration of a component next to the rotational axis
that is caused by the oscillation of the external gears.
Furthermore, there is an unbalanced inertial couple
that is next to the rotational axis of the two external
gears. Because of this, we have incorporated a
counterweight to dampen the vibration from thisinertial couple 5).

3. 1. 2 Test reducer and test conditions
The test sample illustrated in Fig. 5. is comprisedof a hub and a reducer. The hub bearing is lubricatedwith grease while the reducer is lubricated with oil.
The test conditions applied are summarized inTable 2.

Internal pin

2. 2. 2 Motor Design

Externalgear

To reduce the axial size of the unit, we employedaxial gap type motors. Each motor is an SPM motorthat has stators axially opposed to the rotor so that anaxial attraction force is compensated for. The coolingsystem for the motor is comprises of a water-coolingarrangement consisting of a jacket situated at the rearof the stators, and an air-cooling arrangement thatconsists of radiator fins on the outer casing.

Input shaft

Hub bearing
Internalgear

3. Bench test

Oil seal

The entire unit (including the hub, reducer, andmotor section) was subjected to a bench test.

Fig. 5 Test reducer

3. 1 Reducer section
3. 1. 1 Test rig

Table 2 Test condition

Fig. 4 shows a photo of the test rig used. An
induction motor was used as a power source thattransmitted power via a speed changer. An input side
torque meter was used to drive the test reducer. The
output rotation is transmitted through an output side
torque meter and a belt driven reducer to the induction
motor that will create the regenerative braking action.

Input side torque meterMax. input speed

15000min-1

Max. input torque

45Nm

Lubricant oil grade

Industrial lubricant oil ISO VG150

Oil temperature

60〜80˚C

Lubrication system

Oil bath

Oil level (amount of oil) Rotation center of input shaft (at 0 mim-1)

Driving motor

3. 1. 3 Test result
The measurement results for the reducer efficiency
are shown in Fig. 6. The maximum efficiency wasapproximately 95% and the efficiency in vehicle cruise
mode up to 100 km/h is greater than 90%.
Another test was performed using the test conditionin Table 2 with the only change being that the oil level
was lowered by approximately 15 mm. The result of
the second test is illustrated in Fig. 7. A third test was
performed with a different lubricant whose viscosity
grade was VG32 instead of VG150. The result of thistest is shown in Fig. 8. Changing the lubricant amountand viscosity improved the efficiency in both the highspeed and low torque regions. These two changes inlubricant condition proved to have higher efficiency.

Braking motor

Test reducer
Output side torque meter

Fig. 4 Test machine-48-

Development of an In-Wheel Motor Axle Unit

3. 2 Motor
3. 2. 1 Specification

100

The specification of the axial gap motor issummarized in Table 3.
To drive the motor, an inverter was used whose
specification is given in Table 4.

Reducer efficiency %90

2000min-1

80

6000min-1

70

10000min-1
60
50

Table 3 Target specification of axial gap motor

15000min-1
0

10

20

30

40

50

Input torque Nm

Fig.6 Efficiency of reducer

Max. output

20kW

Motor type

Axial gap type, Model SPM

Number of rotor poles

12 polesStator9 slots

Permanent magnet

Nd-based

Cooling system

Water-cooling + air-cooling

Rotary sensor

Hall’s IC

Reducer efficiency %100

Table 4 Specification of inverter

90

80

2000min-1:-15mm
2000min-1:Rotation center
10000min-1:-15mm
10000min-1:Rotation center

70

60

50

15000min-1:-15mm
15000min-1:Rotation center
0

5

10

15

Supply voltage

Max. 450 V

Output30kWDimensions

W400×D500×H248mm

Carrier frequency

20kHz

Drive system

Rectangular wave PWM system

Cooling system

Forced coolingPowering system 120–180˚degrees switchover powering system

20

Input torque Nm

3. 2. 2 Shape Design Based on Magnetic Field
AnalysisFig. 7 Influence of oil level on reducer efficiency

We used a magnetic field analysis technique todesign the shapes of the stators and rotor. Fig. 9shows an example obtained from this analysis where
the magnetic flux density distribution on the stator
core is mapped out. To be able to efficiently generate
torque relative to the level of input current, it isimportant to prevent the saturation of the magnetic
flux density on the stators and rotor. Therefore, in an
attempt to prevent the saturation of the magnetic flux

Reducer efficiency %

100

90

80
2000min-1:VG32
2000min-1:VG150

70

10000min-1:VG32
10000min-1:VG150

60

50
0

15000min-1:VG32
15000min-1:VG150
5

10

15

20

Input torque Nm

Fig. 8 Influence of viscosity on reducer efficiency

Fig. 9 Example of magnetic field analysis-49-

NTN TECHNICAL REVIEW No.75（2007）

density and keep a more compact size motor, wehave redesigned the shapes and dimensions of thecomponents. A prototype motor was fabricated basedon this analysis. Fig. 10 provides a comparison of thedesign analysis values with the actual part test values
in terms of the current vs. torque characteristics of the
motor. The values of the design analysis closely
match those of the actual part measurement result.

4. Actual Vehicle Testing
To verify the concept of our technology and detect
any potential problems of our in-wheel motor type axle
units in practical use, they were mounted a vehicle
and then tested.

4. 1 Test Vehicle Configuration
The unit shown in Fig. 12 was mounted to a vehicle
as shown in the photo in Fig. 13. A few modifications
were made to a current market FF layout vehicle withthe vehicle specifications summarized in Table 5. Our
in-wheel motor type axle units were installed in thetwo rear wheels (with relative ease), so that the initialcharacteristics on an actual vehicle can be easily
obtained.
In order to mount our axle units we modified thetorsion beam suspension, replaced the shock
absorber coil springs, and modified the vehicle-side
mounting sections. Also, we removed the engine andtransmission and installed the driving battery (lithium
ion type) and inverter. We also installed electric
auxiliaries in the engine compartment along side theauxiliary driving battery. In order to cool our axle units,
we installed a water-cooling circuit that included theexisting radiator and an electric water pump.
A schematic of the configured vehicle system canbe seen in Fig. 14. The torque for each wheel iscontrolled separately for the two rear wheels, as
shown in the vehicle information summarized in Table
6.

25
Analysis value
Actual measurement value 2000min-1
Actual measurement value 4000min-1
Actual measurement value 6000min-1

Torque Nm

20
15
10
5
0

0

20

40

60

80

100

120

140

Effective electric current value A

Fig. 10 Characteristic of electric current vs torque3. 2. 3 Efficiency test

Efficiency %For ordinary radial gap motors, the rotor and stators
are made by laminating silicon steel sheets (low coreloss material) in the axial direction to decrease theloss of core material. However, for the rotor of an axial
gap motor it is difficult to laminate the steel sheets thatare used to effectively decrease core loss. The rotor
used in our evaluation testing was made from amaterial other than silicon sheets because a much
simpler fabrication of the rotor was used. The
efficiency measured with this prototype motor is
illustrated in Fig. 11 where the maximum efficiency
was approximately 75%.

Table 5 Specification of test vehicle
Driving system

Two rear wheels

Suspension system

Torsion beam system

Vehicle weight

1350kg100

Driving battery

Lithium ion battery (150V)

80

Brake

Two front wheels only (existing)

Unit cooling system Water-cooling + air-cooling (wind by traveling)

60
40

Table 6 Control of vehicle

2000min-1
4000min-1
6000min-1

20
0
0

2

4

6

8

10

12

14

16

18

20

・Speed on four wheels (vehicle speed)・Motor current
Vehicle information ・Steering angle
・Throttle opening

22

Torque Nm

Axle Unit Control

Fig. 11 Efficiency of motor-50-

Separate Torque Control for Each Side

Development of an In-Wheel Motor Axle Unit

Shock absorber coil spring

Power cable

Axle unit

Torsion beam

Fig.1 2 Unit

Terminal box

Cooling hoseFig. 13 Mounted unit

In-wheel motor unit
Speed (front right wheel)
Steering angle
3-phaseAC current

Battery

Battery

D.C

Inverter

Auxiliaries
Coolant waterpump
Brake negative
pressure pump
Power steering(Right)Rotor rotational angle(Left)
3-phaseAC current

Throttle opening
Speed (front left wheel)

In-wheel motor unit
Fig.14 Structure of vehicle control system

Approximately 250 seconds after the start ofoperation, the temperature increase of the reducer
was 6°C at 20 km/h and 8°C at 40 km/h. Thistemperature rise is relatively insignificant. During thesame running period, the temperature increase of the
motor stators was approximately 15°C at 20 km/h and
30°C at 40 km/h.

4. 2 Vehicle Operating Test
The vehicle was run at speeds up to 40 km/h on a
low friction, paved road that comprised of threedifferent sections: a straight section, a curved section,
and an inclined (hill) section.
To investigate the temperature characteristics of theaxle unit, the test vehicle was run on a chassis
dynamometer at a constant speed of 20 km/h and 40
km/h on a 0% gradient. The temperature wasmeasured on various areas of the axle unit: the motor(stator coils), the reducer (lubricant temperature), andthe motor cooling water. The result of the test is
shown in Fig. 15.
-51-

NTN TECHNICAL REVIEW No.75（2007）

References
70
60

Temperature ˚C1) Ministry of Economy, Trade and Industry, Agency for
Natural Resources and Energy, Gathering on Next
Generation Automobiles and Fuels: Next Generation
Automobiles and Fuels Initiative Summary (2007).
2) Yasuki Tahara, Ryoji Mizutani, Yuki Tojima, Masafumi
Sakuma: Development of In-Wheel Motor System,
Society of Automotive Engineers of Japan, Lecture
Session Preprints, No.131-06, 20065703 (2006)
3) Makoto Kamachi et al.: Improvement of Vehicle
Dynamic Performance by Means of In-Wheel Electric
Motors, Mitsubishi Motors Technical Review, No.18(2006) 107-113
4) Rio S. Zhou, Fukuo Hashimoto: Highly Compact
Electric Drive for Automotive Applications, SAE paper,
2006-01-3037
5) Minoru Suzuki, Dawei Wang: NTN Technical Review,
No.73 pp.56-59 (2005)6) Muneharu Morozumi: Theory and design calculation
technique for planetary gears and differential gears,
Nikkan Kogyo Shinbun, Ltd., pp.1-6 (1989)

40km/h Motor
40km/h Reducer
40km/h Cooling water
20km/h Motor
20km/h Reducer
20km/h Cooling water

50

40
30

20

10

0

100

200

300

400

500

600

Run time sec

Fig. 15 Temperature characteristic of unit

5. Conclusion
Our cycloid reducer mechanism was found to haveexcellent power transmission performance (amaximum output of 20 kW, maximum input speed of
15000 min-1, and maximum torque of 45 Nm). As well
as excellent performance, this mechanism has amaximum efficiency of approximately 95%. For the
motor section, we fabricated an axial gap motor andan inverter and then evaluated the characteristics ofthe motor. We mounted prototype axle units to avehicle, each comprised of a reducer and an electricmotor. We then tested the vehicle under constantvelocity conditions on a chassis dynamometer. Lastly,
we ran the vehicle along a paved road on straight andcurved sections. Based on this vehicle test, we haveverified trouble-free operation of our axle units.

Photos of authors

Minoru SUZUKI

Kayo SAKAI

Koichi OKADA

Yusuke MAKINO

New Product Development Dept.
New Product Development R&D
Center

New Product Development Dept.
New Product Development R&D
CenterMechatronics Research Dept.
New Product Development
R&D Center

Mechatronics Research Dept.
New Product Development
R&D Center-52-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Paper ]Development of an Electromechanical Brake

Tatsuya YAMASAKI**
Masaaki EGUCHI**
Yusuke MAKINO**The NTN electromechanical brake (EMB) system wasdesigned and launched with the intention of improving overall
performance of future automotive braking systems. NTN has
designed and developed a new linear actuator that can beapplied to this system. This paper reports on the configuration,
design principles, efficiency calculation methods and related
experiments.

is used as a parking brake without any modification, aseparate mechanism will be needed to lock therevolution motion of the motor because a linear
actuator such as a ball-screw lacks a thrust holding
function. To address this issue and satisfy thefunctions required for such an electromechanicalbrake unit (high load conversion ratio, fretting-resistantquality, load-holding function), NTN has invented aunique linear actuator mechanism that doesn’t requirean accompanying mechanism and developed aunique electromechanical brake unit that includes anelectric motor.
In this paper, we describe the constitution andoperating principle of this electromechanical brake
unit, the method for calculating the efficiency of thebrake unit, as well as the evaluation tests we have
performed.

1. Introduction
In the automotive industry there has been an
increased emphasis on vehicle safety. Improvement inbrake technology has greatly contributed to stablerunning of vehicles. Increased functionality has
resulted in products like ABS, ESC, and brake assist1)-2). An example of the increased functionality of
automotive brakes is improvement in control
techniques for hydraulic brakes. Furthermore, in aneffort to continue this improvement in functionality andreduction in environmental impact, automotive
components manufacturers and car manufacturers aredeveloping electromechanical brake (EMB) systems.
Many of newly developed electromechanical brakesystems employ linear actuators such as ball-screws
and ball-ramps (torque cams). However, when any ofthese linear actuators is used to develop the sufficientthrust required to brake a traveling vehicle, greater
input torque is needed because the load conversion
ratio with the linear actuator alone is insufficient. To
design a more compact, lightweight electromechanicalbrake unit, the motor must be more compact.
Therefore an independent reducer needs to beincorporated. Additionally, the electromechanicalbrake is situated in the “unsprung” section of thevehicle and will be subjected to violent vibration.
Therefore it must be positively fretting-resistant.
Incidentally, when the electromechanical brake unit

Legends
a，b
DS
DP
dBS
dO
F
k
LP

: half width of Hertz contact ellipse on
threaded surface: sun roller outside dia.
: planetary roller outside dia.
: ball center diameter on ball-screw: outer ring inside dia.
: force acting on the piston: number of effective thread ridges on eachplanetary roller = lP/P
: thread lead on planetary roller**New Product Development Dept. New Product Development R&D Center**Mechatronics Research Dept. New Product Development R&D Center-53-

NTN TECHNICAL REVIEW No.75（2007）

LO
lP
n
p
PSC

: thread lead on outer ring
: effective length of threaded portion on eachplanetary roller: number of planetary rollers: thread pitchF: load acting on one screw = ―――――
nk cosβPN: radial circumferential load from one planetaryFroller, acting on sun roller = ――・tanβnPr: circumferential tangential force from one

2. Linear actuator
2. 1 Constituent Elements and Operating Principle
The constitution and operating principle of NTN’slinear actuator are discussed below.
NTN’s linear actuator comprises, as shown in Fig.
1, a sun roller, planetary rollers, an outer ring, acarrier, a piston and a threaded member whichconstitutes the screw of the outer ring. Several
planetary rollers are shrink-fitted between the sunroller and the fixed outer ring; thereby the planetaryrollers rotate and revolve as the sun roller rotates.
Spiral threading is provided on the outercircumference of each of the planetary rollers as wellas on the inner circumference of the outer ring,wherein these threads have an identical pitch so thatthey can be correctly meshed with each other.
As a result of this layout, while revolving androtating, the planetary rollers will be axially shifted
relative to the outer ring. The carrier and piston eachsupport rotation and revolution of the planetary rollers;then, the rotary motion of the sun roller is finallyconverted into the linear motion of the piston.
Suppose that there is no slipping at the contact
points between the outer circumference of sun roller,
outer circumferences of planetary rollers and inner
circumference of outer ring, the amount of axial
displacement x of the piston relative to the rotary
motion of the sun roller can be expressed with theformula (1). At the same time, the amount of axial
displacement xBS of the nut on the ball-screw can be
expressed with the formula (2). The axial
displacement x of the piston is dependent on the
difference between the lead angle of the thread on theplanetary rollers and that on the outer ring; thereby theaxial displacement x of the piston (our linear actuator)
can be smaller than the axial displacement xBS of thenut on the ball-screw. In other words, compared with avalue obtained with the ball-screw arrangement, theload conversion ratio, which can be defined as thethrust relative to the input torque, can be greater with
our linear actuator.

2TS
planetary roller, acting on sun roller = ＝―――nDS
: axial tangential force acting on sun roller: sun roller torque: axial displacement of piston: axial displacement of nut (threaded shaft) on
ball-screw
WIN : input (workload)
WOUT : output (workload)
WREV : loss from frictional torque occurring on
revolution supporting bearing
WROT : loss from frictional torque occurring on
rotation supporting bearingWS: loss from frictional torque occurring on sunroller supporting bearing
WSC : loss from slipping on screw threadα: equivalent lead angle
αBS : thread lead angle on ball-screw
αP : thread lead angle on planetary roller
αO : thread lead angle on outer ringβ
: thread flank angle on planetary rollerδ: amount of offset at contact point on screw
threadη: efficiencyμ: friction coefficient on circumferential contact
surfaceμSC : friction coefficient on screw threadθBS : angular displacement of threaded shaft (nut)on ball-screwθS: angular displacement on sun roller
θREV : angular displacement of revolution onplanetary roller＝［DS（dO＋DS）］
・θS
θROT : angular displacement of rotation onplanetary roller＝［−DS（dO−DS）］
・θS
θ'ROT : angular displacement of rotation on planetaryroller (relative to carrier)＝θROT−θREV
Pa
TS
x
xBS

dO
x＝――・
（tanαO−tanαP）
・θREV …………(1)2dBS
xBS＝――・tanαBS・θBS………………………(2)2Also, we have attempted to improve fretting
resistance of our electromechanical brake by applying
a negative clearance preload in the radial direction
among the sun roller, planetary rollers and outer ring.
To achieve, at a lower cost, both adjustment of radial
clearance and formation of thread, the thread is-54-

Development of an Electromechanical Brake

Threaded member

Forwarddisplacement

By referring to Fig. 2, suppose that the torque being
input from the sun roller causes the piston to moveaxially. Then, the efficiency of the linear actuatorηandthe relation between the work input and work output
can be expressed with the formulas (3) and (4),respectively.

Planetary roller

Rotation

WOUT
F・x
η＝―――＝――― …………………………(3)
WIN
TS・
θS

Sun roller
PistonCarrier
Outer ring

WOUT＝WIN−WREV−WROT−WS−WSC ………(4)Fig. 1 Schematic of linear actuator

The torque loss from each bearing was calculated
utilizing the Palmgren 3). experimental formula for
friction torque.
Now, we would like to discuss the load acting on thesun roller supporting bearing. The normal force andtangential force from a planetary roller, shown in Fig.
3, act on the sun roller. From the relation with africtional circle, we determined the axial tangential
force Pa with the formula (5) below, and took theresultant force value as the axial load acting on thesun roller supporting bearing.

formed on the outer ring by adding a coil-shaped
thread member in the spiral grooves formed on thebore surface of the outer ring.

2. 2 Efficiency
In order to be able to determine the efficiency of ourlinear actuator, we provided the following assumptions[1] through [4]:[1] The thrust occurring on the piston is uniformly
carried by all the planetary rollers and threads.
[2] Because of the smaller lead angle (not greater than
5 degrees), the circumferential component of theload occurring on the thread can be ignored.
[3] No slipping occurs at the contact points between
the outer circumference of the sun roller and theouter circumference of each planetary roller andbetween the inner circumference of the outer ringand the outer circumference of each planetary roller.
[4] The areas where loss appears to occur are therevolution supporting bearing, rotation-supporting
bearing, and sun roller supporting bearing andcontact surfaces on threads.
In addition to these areas, the axial sliding at thecontact areas between the planetary rollers and sunroller and between the planetary rollers and outer ringis regarded as a loss. However, this sliding
accompanies rolling contact, and is considered to be
fairly small.
Rotation supporting bearing2
Pa＝n √（μPN）−Pr2 ………………………(5)

PN

Pa/n

Fig. 3 Load on sun roller

Next, we would like to explain the losses arisingfrom sliding on the screw thread. On the screw
threads between the planetary rollers and outer ring,the Hertz contact ellipse shifts, as illustrated in Fig. 4,
while the planetary rollers rotate. Suppose that eachplanetary roller moves along the inner circumferenceof the outer ring. Then, the contact ellipse on thescrew thread shifts with the sliding motion thatconsists of the tangential displacement U of the centerof ellipse (expressed with the formula (6)) and therotational displacement θ on the center of ellipse(expressed with the formula (7)).

Housing

Revolution
supporting
bearing

Thrust：F
Axial
displacement：x

Sun roller
supporting bearing

Pr＝2TS/nDS
μPN

dO
DP
U＝ ――−δ ・θREV＋ ――−δ ・（θROT−θREV）
2
2（

Torque：TS
Angulardisplacement：θS

）

（

）

＝δθROT ……………………………………(6)θ＝（θROT−θREV）
・COSβ＝θ'ROT・COSβ ……(7)Fig. 2 Relation of input and output-55-

NTN TECHNICAL REVIEW No.75（2007）

Then, take the sliding-induced losses within thecontact ellipse as the sum of losses within a miniaturearea such as a one given in Fig. 5. Then, from theamount of sliding V (x, y), determined with the formula(8) and the normal load f (x, y), determined with theformula (9), the sliding-induced loss WSC on the screw
thread can be expressed with the formula (10).

Under the calculation conditions given in Table 1,
the thrust F and efficiencyηfor the units No. 1 through4 were calculated, and the result is plotted in Fig. 6;
and the loss ratios relative to the inputs to the unit No.
1 are illustrated in Fig. 7. For these calculations, it
was assumed that μ=μSC = 0.1. From Fig. 6, itshould be understood that the thread can be variously
specified and a variety of linear actuators eachfeaturing unique efficiency can be designed. With Fig.
7, it is possible that the proportion of the loss resulting
from sliding on the screw thread is large andincreases as the torque increases.

V(x, y)＝ (U＋rθsinφ)2＋r 2θ2 cos2φ ……… (8)f (x, y)＝

3PSC
2πab

1−

x2 y2
−
・dx・dy …… (9)a2 b2

WSC＝nk∫μSC・f(x, y)・V(x, y)…………… (10)Table 1 Calculation condition
Outer ringContact ellipse

Technical data

No.1

No.2

No.3

No.4

DS, mm

12

←

←

←

DP, mm

30

←

←

←

dO, mm

72

←

←

←

n

3

←

←

←

LP, mm2

2

0

3

LO, mm

4

2

2

6

p, mm

2

←

←

3

IP, mm

16

←

←

←

δ, mm

0.3

←

←

←

δ

Screw thread

Planetary roller

40

80

30

60

20

40

dx

x

U
dy
r
φ
V(x.y)

Efficiency

20

No. 1
No. 2
No. 3
No. 4

y
0

rθ

Thrust

10

0

1

0
2

3

Torque Nm

Fig. 6 Calculation results of thrust and efficiency

Contact ellipse

Fig. 5 Contact ellipse

Loss ratio %80
Revolution supporting bearing
Rotation supporting bearing
Sun roller supporting bearing
Screw thread

60

40

20

0

0

1

2

Torque Nm

Fig. 7 Ratio of losses

-56-

3

Efficiency %Thrust kN

Fig. 4 Contact ellipse in thread of the screw

Development of an Electromechanical Brake

3. 2 Efficiency and load holding function

3. Evaluation for characteristics oflinear actuators

The test results of thrust and efficiency for the units
A and B are given in Fig. 9. These results coincidewith the result of simulated calculation. This means
our calculation practice is reliable.
The relation between the torque and thrust on thelinear actuator in one cycle (a duration where thethrust is increased from a zero level to a particularlevel, and then is decreased to a zero level) is shownin Fig. 10. When the thrust is decreased, a negative
torque is necessary on the unit A while a positive
torque occurs on the unit B. This is because the sunroller rotates by the thrust and, consequently, its thrust
holding function is lost.
To address this issue, we developed the relationshipbetween the equivalent lead angle of the linear
actuator defined with the formula (11), efficiency of thelinear actuator when it develops 30 kN and availabilityof its thrust holding function, and this relation is plottedin Fig. 11. With a greater equivalent lead angle, theefficiency is better; when the equivalent lead angle is0.331 deg or greater, the thrust holding function is notavailable. To sum up, it must be understood that thelinear actuator needs to be designed to have an
equivalent lead angle of approximately 0.3 deg so thatit can have the thrust holding function.

3. 1 Units being tested and test method
We have evaluated the efficiency and thrust holding
function of our linear actuator. The specifications of
the linear actuator units tested are summarized inTable 2, and the configuration of the test rig is
illustrated in Fig. 8. With the test rig, torque wasapplied to the linear actuator by, for example, applying
a weight to the lever, thereby the efficiency of ourlinear actuator was determined based on the torqueinput to the sun roller and the thrust from the piston
detected by the load cell.

Table 2 Specifications of linear actuator
Unit

A

B

C

D

E

F

DS, mm

12

8

15

←

←

←

DP, mm

30

26

28.5

←

←

←

dO, mm

72

60

72

←

←

←

n

3

←

4

←

←

←

LP, mm

2

0

2

0

2

3

LO, mm

4

2

4

2

2

6

p, mm

2

←

←

←

←

3

α, deg

0.173

0.536

0.220

0.419

0.640

0.331

Load cell

Test unit

2xα＝tan-1 ――――
DS・θS

（

） ………………………(11)Torque meter

Lever

40

80

30

60

30

60

20

40

20

40

20

10

20

10
Thrust Efficiency

Thrust Efficiency

Test
calculation

Test
calculation

0
0

1

2

0

0

0

3

0

Torque Nm(a)Unit A

1

2

Torque Nm(b)Unit B

Fig. 9 Test results of thrust and efficiency

-57-

3

Efficiency %80

Thrust kN

40

Efficiency %Thrust kN

Fig. 8 Test rig for measuring efficiency

40

40

30

30

Thrust kN

Thrust kN

NTN TECHNICAL REVIEW No.75（2007）

20

20

10

-1

10

0

1

2

3

4

-1

0

1

2

Torque Nm

Torque Nm(a)Unit A(b)Unit B

3

4Fig. 10 Relation between torque and thrust

4. Evaluation for electromechanicalbrake units

80

4. 1 Electromechanical brake units tested

Efficiency %60

We have manufactured a prototype
electromechanical brake unit that can be incorporated
into an actual car. It has been evaluated on a test rig
and tested as mounted on an actual car. The
specifications of the electromechanical brake units
and motor used are summarized in Tables 3 and 4,
while the configuration of the electromechanical brake(EMB) unit is illustrated in Fig. 12 and that of the
motor is given in Fig. 13. So that the axial length of
EMB unit can be shorter, the motor is arranged
parallel with the linear actuator and the motor itself is
an axial clearance-type motor that has two stators.
From this motor, the torque is transmitted via a gearing
to the sun roller. For convenience of incorporation into
an actual car, the linear actuator used is the type B unit
described in Sec. 3, which features a compact size.

40

20
Load holding function is valid.
Load holding function is not valid.

0
0

0.2

0.4

0.6

0.8

Equivalent lead angle deg.

Fig. 11 Relationship between equivalent lead angle,
efficiency and load holding

Pad

Y

Linear actuator

X

Motor

Y

X

SECTION X-X

SECTION Y-Y
Brake rotor

Fig. 12 Prototype EMB unit-58-

Development of an Electromechanical Brake

Table 3 Specifications of EMB units
Characteristic

Rotor

StatorSpecification

Max. thrust

30kN

Size (excluding caliper)

150mm×82mm×85mm

Mass (excluding caliper)

3.8kg

Gear

Table 4 Specifications of motor
Characteristic

Specification

Max. output

140W

Max. torque2NmMax. speed

1500min-1

Y

Fig. 13 Axial gap motor

4. 2 Evaluation about response
This is because with a greater voltage applied, themaximum motor speed is greater.
Next, we want to discuss the effect of an initialclearance between the piston and load cell. Let us
think of the cases where the initial clearance is set to
0.2 mm. When the voltage applied is 12 V, the timeneeded for a thrust to occur takes approximately 0.1
second while this time span is as short asapproximately 0.05 second when the voltage is 20 V.
Thus, the size of initial clearance significantly affects
the response time.

On the test rig illustrated in Fig. 14, we evaluatedthe response time relative to thrust on our EMB unit
(increase and decrease characteristics). The thrust
was measured with a load cell that was mounted inplace of a brake rotor and pad. The test conditions are
summarized in Table 5, and the test results areplotted in Figs. 15 and 16.
In our test project, a particular voltage was fed into
the motor 0.1 sec. after the beginning of eachmeasuring run, thereby the thrust increase/decreasecharacteristics were evaluated. Under any
combination of test conditions, the time-dependent
variation in the thrust is nearly constant in the region
approximately 10 kN or greater. In the thrust range 10
kN or lower, this variation does notappear to be linear. It is possible, by applying agreater voltage to the motor, to promote the variationin thrust, in either a thrust increase or decrease mode,
to shorten the time needed to reach a targeted load.

Table 5 Test condition
Parameters

ThrustIncrease characteristic

Voltage applied
Initial clearance―

0mm，0.2mm

PistonLoad cell

Linear actuator

Fig. 14 Test rig for measuring the response time

-59-

Decrease characteristic

12V，20V

Motor

3000
Thrust
Speed
Current
Voltage applied

40

2000

30
20

1000

10
0
0

0.1

0.2

0.3

0.4

0.5

0
0.6

3000
Thrust
Speed
Current
Voltage applied

50
40

2000

30
20

1000

10
0
0

Thrust kN, Current A, Voltage applied V

60

3000
Thrust
Speed
Current
Voltage applied

50
40

2000

30
20

1000

10
0
0

0.1

0.2

0.3

0.4

0.5

0.2

0.3

0.4

0.5

0
0.6

Time sec(b) Voltage applied 20V, clearance 0mm

Speed min-1

Thrust kN, Current A, Voltage applied VTime sec(a) Voltage applied 12V, clearance 0mm

0.1

0
0.6

60

3000
Thrust
Speed
Current
Voltage applied

50
40

2000

30
20

1000

Speed min-1

50

60

Speed min-1

Thrust kN, Current A, Voltage applied V

60

Speed min-1

Thrust kN, Current A, Voltage applied V

NTN TECHNICAL REVIEW No.75（2007）10
00Time sec(c) Voltage applied 12V, clearance 0.2mm

0.1

0.2

0.3

0.4

0.5

0
0.6

Time sec(d) Voltage applied 20V, clearance 0.2mm

3000
Thrust
Speed
Current
Voltage applied

50
40

2000

30
20

1000

10
0
0

0.1

0.2

0.3

0
0.4

60

3000
Thrust
Speed
Current
Voltage applied

50
40

2000

30
20

1000

Speed min-1

Thrust kN, Current A, Voltage applied V

60

Speed min-1

Thrust kN, Current A, Voltage applied V

Fig. 15 Response time (increase characteristic)

10
0
0

0.1

0.2

0.3

0.4

Time sec

Time sec(a) Voltage applied 12V(b) Voltage applied 20V

0.5

0
0.6

Fig. 16 Response time (decrease characteristic)

configuration shown in Fig. 18, a thrust value detectedby the strain gage relative to a command value fromthe controller is fed back to the controller in order tocontrol the motors. We attempted to brake the car
traveling at a speed of approximately 30 km/h, by only
actuating the EMB units on the rear wheels, and could
reliably stop the car. For this test, we used a 12-V
battery for automotive applications.

4.3 Test on actual car
As illustrated in Fig. 17, we incorporated our
prototype EMB unit into each of rear wheels on thetest car. Our EMB unit could be readily installed to thetest car only with adjustment to the mounting position,
and without modification to the area including the
knuckle. Also, we have already verified that our EMB
unit can be incorporated into the front wheels of the
car without any significant modification. On the system-60-

Development of an Electromechanical Brake(EMB) unit that includes our linear actuator and anelectric motor, and obtained the following results:(1) We have developed an efficiency calculation
technique for our linear actuator, and verified thevalidity of this technique through a series of tests.
(2) We have measured the thrust and efficiency of ourlinear actuator, thereby we have clarified thethread specifications for satisfying the thrust
holding function.
(3) We have learned the characteristics of the EMB
unit (a unit including an electric motor) through a
response evaluation test and a braking test on an
actual car.

Fig. 17 Installation of the unit

References
1) Tomohiko ADACHI: The Electronic Stability Control
System, Journal of Society of Automotive Engineers ofJapan, Vol.60, No.12 pp.28-33 (2006)
2) Bo Cheng, Tetsuo Taniguchi, Tadashi Hatano, Toshiya
Hirose: Effect of Brake Assistance System in
Emergency Situation, Society of Automotive Engineersof Japan, Lecture Session Preprints, No.20065894
3) A. Palmgren：Ball and Roller Bearing Engineering,
3rd ed., Burbank, Philadelphia, (1959) 34.

5. Conclusion
NTN has invented a unique linear actuator for EMB
units that will be marketed in the near future, wherein
our linear actuator, though not having any additional
mechanism, not only boasts a high level of load
conversion ratio and high degree of fretting resistancebut also features a unique load holding function.
Then, we have evaluated an electromechanical brake

Motor driving battery

Hydraulic brake

Strain gage

Electromechanical brake unit

CommandInverter

Motor
phase

Inverter driving
battery
Inverter

Controller

Command

Electromechanical brake unit
Motor driving battery

Hydraulic brake

Fig. 18 Prototype EMB system diagram

Photos of authors

Tatsuya YAMASAKI

Masaaki EGUCHI

Yusuke MAKINO

New Product Development Dept.
New Product Development R&D
Center

New Product Development Dept.
New Product Development R&D
CenterMechatronics Research Dept.
New Product Development R&D
Center-61-

AC（U, V, W）

Strain gage

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Engine Part Technical Trends
and New Engine Products

Yoshiaki RYOUNO*
Kouichi ONIMARU*

Kenichi KAWABATA**
Shinji OOISHI**

Engine specifications have been influenced by environmental andsocial conditions. Now, fuel-efficient engines are expected to reduce
the emission of CO2, a contributor to global warming. In particular,
engine parts with extended lifespans, reduced frictional losses, andreduced weight are needed to meet the new 2015 fuel-efficiency
standard targets. It goes without saying that all parts must be
durable and reliable. This report discusses new engine partproducts, developed by NTN that, fulfill these requirements.

automakers are committed to improvements in power
performance of car engines, fuel economy andemissions performance, and in this context, they havebeen achieving novel technological developments. To
help commercialize these new automotive
technologies, NTN has developed a diverse group of
new mechanical elements.

1. Introduction
EU released its new environmental regulation
stipulating that by the end of 2012, the CO2 emissions
from new cars sold in EU have to be reduced to an
average of 130 g/km; accomplished throughimprovement in engine performance. Also in the USA,
there is a fuel efficiency improvement program
according to which gasoline consumption has to be
reduced by 20% in 10 years beginning in the year
2007.
In Japan, a regulation will be introduced, which will
impose new fuel-efficiency standards (see Table 1) forsmall trucks, small buses and passenger cars andthese standards must be satisfied by the end of theyear 2015. In particular in terms of the fuel efficiency
with passenger cars, the projected year 2015 level willbe 23.5% better compared with the actual year 2004
level, and will be 29.2% better compared with thecurrent fuel efficiency level (the target for FY2010).
Under such demanding circumstances, the

2. Power loss factors on cars--friction
losses on automotive engines andrequirements for NTN products forautomotive engines
2. 1 Power loss factors on cars
Typical factors contributing to power loss on cars
are summarized in Fig. 1.
The engine friction accounts for 27% of the entire
power loss on an average car: therefore, we believethat a decrease in friction on an automotive engine will
help improve the fuel efficiency of the car. We also
believe that a lighter weight car will contributes to an
improvement of fuel efficiency of the car.

Table 1 2015 year fuel-efficient target(improvement rate compare with 2004)1）
Car category

Actual year 2004 Projected year 2015
fuel efficiency
fuel efficiency

2. 2 Friction losses on automotive engines
A typical breakdown of friction ratios with anautomotive engine is illustrated in Fig. 2.
[1] The major friction losses on an automotive engine
are derived from the valve system, crankshaft, oilpump system, valve drive system, and piston
system.

Improvement in fuelefficiency overyear 2004 level

Passenger car 13.6（km/R） 16.8（km/R）

23.5%

Small bus

8.3（km/R）

8.9（km/R）

7.2%

Small truck

13.5（km/R） 15.2
（km/R）

12.6%**Automotive Engineering Dept. Automotive Sales Headquarters**Needle Roller Bearing Engineering Dept. Automotive Sales Headquarters-62-

Engine Part Technical Trends and New Engine Products

Friction on
torque converter
4%
Frictionon A/T
9%

Piston system
Oil pump system
Valve drive system

Crankshaft system
Valve system

Friction loss ratio

100％

Friction on engine
27%
Air drag
19%

32

33

5
7

18

20

4

37

20

50％

3
28

Friction fromrolling tires
21%

32
36

Weight of car
20%

17
0％

1000

3500

8
6000

Engine speed min-1

Fig. 1 Friction loss ratio at the car (HWY mode)2）

Fig. 2 Friction loss ratio of the engine3）

The proportions of friction loss ratios can varyfrom rotation to rotation of the engine.
Notwithstanding, reduction in these friction losses
will help improve the fuel economy of the car.
(Our present development program for the
products intended for automotive engines does notcover the parts associated with pistons.)[2] The friction loss derived from the oil pump is highat 3500 min-1 and 6000 min-1. There is a clear
interrelation between the discharge amount andfriction loss of the oil pump: by reducing thenecessary discharge amount, the friction loss ofthe oil pump. By reducing the necessary discharge
amount, the friction loss on of the oil pump can be
decreased, and the necessary pump size can besmaller. Therefore, to reduce fuel consumption,
reduction in the amount of oil fed by an oil pumpappears to be important.

engines are summarized in Table 2. By reducing thefriction losses with various parts for engines in order toreduce fuel consumption, we will be able to reduce
CO2 emissions and conserve energy resources.

2. 4 Newly developed NTN products forautomotive engines
The newly developed products for automotiveengines described in this paper are illustrated in Fig.
3.
Among them, the following products are describedin the sections below:(1) Oil-less chain tensioner(2) Rocker arm FA needle roller bearing(3) Low-friction capable needle roller bearings (forsupporting crankshaft, for supporting
camshaft)(4) Ball-screw unit for valve continuous variablemechanism
See the associated "New Product" section.
(5) End-pivot type mechanical lash adjuster
See the associated "Technical Paper" section.

2. 3 Requirements for NTN products forautomotive engines
The requirements for NTN products for automotive

Table 2 Requirements for engine products
Portion on engine

Products for engines and requirements

Valve system• Ball-screw unit for valve continuous variable mechanism(reduced pump loss)• FA needle roller bearing (smaller size, longer life)• End-pivot type mechanical lash adjuster (smaller oil pump)

Crank system• Crankshaft supporting split needle roller bearing(lower frictional torque)

Valve drive system• Camshaft supporting needle roller bearing (lower frictional torque)• Oilless chain tensioner (smaller oil pump)Engine main body

• Each component (smaller size, lighter weight)-63-

Objective

Effects on the environment

Reduction in CO2 emissions(prevention of global warming)
Lower fuelconsumption

Conservation of energy resources(wise utilization of resources)

NTN TECHNICAL REVIEW No.75（2007）

Rocker arm
FA needle roller bearing

Camshaft supporting
needle roller bearing

Ball-screw unit for valve continuous
variable mechanism

Requirements: Smaller size, longer life

Requirements: Lower frictional torqueRequirements: Lower pump loss
(See the relevant product introduction section.)

Valve continuous
variable mechanism 4)

End-pivot typemechanical lash adjuster

Crankshaft supporting
split needle roller bearing

Requirements: Smaller oil pump
(See the relevant technical report section.)

Oilless chain tensioner
Requirements: Smaller oil pump

Requirements:
Lower frictional torqueFig. 3 Typical products for engine

NTN has been mass-producing two types of chain
tensioners; -- "buttress thread type" and "ring type", -each using the oil supplied from the engine as adamping means. To satisfy the current market need,
we have developed an "oil-less type" chain tensioner
that does not need oil supplied from the engine.
Our oil-less chain tensioner is described below.

3. Oil-less chain tensioner
Recently, the performance of automotive engines
has been improving; and with this an increasing
number of hydraulically actuated systems areincorporated into automotive engines.
As a result, oil pumps of larger size are used andthe oil piping system is increasingly complicated. To
address environmental issues, however, needs are forsmaller, lighter oil pumps and reduced consumablelubrication needs.

3. 1 Features
There are two representative functions required forthe chain tensioner. They are the "damper mechanism"-64-

Engine Part Technical Trends and New Engine Products

gradually delivered to the outside through the leakgap defined by the cylinder and plunger. This
causes the hydraulic damper to trip. Incidentally,
depending on the engine characteristics, a relief
valve may be integrated which releases oil to theoutside when the internal pressure in the pressurechamber exceeds a particular threshold level.

and the "no-back mechanism".

[1] Damper mechanism
This mechanism is necessary to maintain constanttension on the timing chain. In an over -- tensioned
situation, the life of timing chain will be shorter and/or
friction loss can occur on the timing chain. Not enoughtension can cause the chain to wobble, possibly
leading to excessive noise and teeth being skipped.

The structure of our oil-less chain tensioner isillustrated in Fig. 5.
[1] When the chain has slack, the rod moves
according to the force of the return spring, andfollows the behavior of the chain. The buttressthread provides a mechanical damping force sothat the chain is adequately tensioned.
[2] When the chain is tensioned, the buttress thread is
dynamically forced in from the operation of thelever guide and rod which protects the chain fromover-tensioning.
[3] The buttress thread is independent of the rod, sothat the effect of rotating force on the buttressthread resulting from the sliding action relative tothe lever guide is eliminated.
[4] The oil prefilled in the tensioner lubricates thebuttress thread. This arrangement means that thetensioner is not affected by the engine oil (type,viscosity, pressure, deterioration), and can perform
stably.

[2] No-back mechanism
While the engine is at a standstill, and depending ona phase relation between the camshaft and valve, aload may be applied to the tensioner even after theengine is turned off. When the engine is restarted inthis position, the amplitude on the tensioner can beexcessively large being in the forced-in position, which
can result in noise. To solve this problem, a “no-backmechanism” will be integrated in the chain tensioner toprevent the plunger from ending up at the forced-inmode while the engine is at a standstill.
To provide a damper mechanism, NTN has adopted
hydraulic damper types that utilize oil from the engine;
namely, the buttress and ring types.
Our “no-back mechanism” employs a buttressthread structure (buttress thread type) or a wedge
structure (ring type) consisting of a ring; causing theplunger to be forced in.
The damper mechanism on the oil-less chain
tensioner consists of a buttress thread and a spring
damper; while. The “no-back mechanism” on thetensioner relies on the frictional resistance of thebuttress thread.
In summary, since our oil-less chain tensioner does
not need oil supplied from the engine, the oil pump
can be smaller and requires no oil piping system orpipe connections on the side of the engine.
Consequently, we can expect a lighter engine
arrangement, lower fuel consumption, and a lessexpensive engine system.

3. 3 Evaluation of performance
The motor characteristics of an oil-less chain
tensioner on an ordinary reciprocating engine areillustrated in Fig. 6.
With specification A, the chain tensioner has a relief
structure, and, therefore, runs stably over the entire
range of running speeds, with minimal variation inchain tension. By adequately adjusting the relief
setting, this chain tensioner can be used on engineswith diverse running characteristics.
With specification B, the chain tensioner lacks arelief structure, and, as a result, the chain tension ishigh at higher running speeds. The drawback of this isthat, the amplitude of the tensioner is very small,
making this specification suitable for an engine thatfeatures smaller load variation.

3. 2 Construction and operation
Structures of the buttress thread and ring type chain
tensioners are illustrated in Fig. 4.
[1] When the chain has slack, the plunger moves
according to the force of the return spring, causing
the chain to be adequately tensioned.
[2] During this movement, the internal pressure of thepressure chamber drops, and the check valve
opens allowing the oil to be supplied from theengine side.
[3] While the chain is tensioned and the plunger isforced in, the oil in the pressure chamber is-65-

NTN TECHNICAL REVIEW No.75（2007）

Relief valve

Screw rod

Return spring A

Return spring B

Return spring

CylinderPlunger

Plunger

Relief valve

Check valve

Oil fill port
Buttress Cylinder
threadOil fill
port
Check valve

Ring

Ring type

Buttress thread typeCam sprocket

Fig. 4 Structure of buttress thread type/ring type chain tensioner 5）6）

[Rapid response]
Return spring

Chain tensioner[Reliable sealing]
Oil seal[Lighter weight]
Aluminum cylinder

Chain

Rod

Lever guide
Chain guide
Crank sprocket[Spring damper]
Prevention of
overtensioned chain[Prevention of deviation of rod]
Two-point support with wear ring
[Step-less no-back mechanism]
Buttress thread structure

Specification A: w/ spring damper

Specification B: w/o spring damper
Fig. 5 Structure of chain tensioner without oil supply

Effective tension N

Chain tension

0

1000

2000

3000

4000

5000

6000

7000

1200
Slack side
Tensioned side

800
400
0
-400
-800
-1200

0

1000

2000

Engine speed min-1
2
1.5
1
0.5
0
-0.5
-1
-1.5
-2

0

1000

2000

3000

4000

3000

4000

5000

6000

7000

5000

6000

7000

Engine speed min-1
Effective tension N

Direction for chain tensioning→
Direction for chain tensioning→

Plunger displacement mm Plunger displacement mm

Specification B

Specification A

Tensioner deflection
2
1.5
1
0.5
0
-0.5
-1
-1.5
-2

5000

6000

7000

1200

Slack side
Tensioned side

800
400
0
-400
-800
-1200
0

1000

2000

3000

4000

Engine speed min-1

Engine speed min-1

Fig. 6 Motoring characteristics of chain tensioner without oil supply-66-

Engine Part Technical Trends and New Engine Products

rollers and shaft-each made of bearing steel.
[2] The austenite grain boundaries on a conventional
standard heat-treated bearing steel material andthose of an FA-treated bearing steel material are
shown in Photo 2.
The size of the crystals from the FA-treatment
are up to the size of those of standard heat-treatedbearing steel.

4. Rocker arm FA needle roller bearing
There is a steel making technology featuring finer
crystal grain size for a given steel material in order toimprove the fatigue strength. To use this advantage,
NTN has developed a special heat treatment process(FA-treatment), wherein the crystal gain size on the
bearing steel has been reduced to 1/2 or smaller
compared with conventional bearing steel materials.
Because of this novel technology, the FA needle
roller bearing for an engine rocker arm boasts a life atleast 3.7 times as long as that of conventional rocker
arm bearings. Consequently, for given operating
conditions, the width and mass of our needle roller
bearing can be as small as approximately 75% thesize of a conventional needle roller bearings for thesame application.

4. 2 Performance evaluation
Comparisons between FA-treated bearings andconventional bearings, in terms of bearing life, areprovided in Fig. 7. Comparisons between FA-treatedbearings and conventional bearings, in terms ofbearing life deriving from differences in face pressures
are shown in Fig. 8.
Under a given set of operating conditions, the life of
FA-treated bearings is 3.7 times greater than that ofconventional bearings; The face pressure on therollers and bearing with the FA-treated material can be
13% greater compared with those made with standard
heat treatments. This advantage allows for alightweight, compact needle roller bearing design.
A rocker arm bearing may be attached to the armwith one of two methods. One method is staking thearm with the shaft. The second method is to press-fit
the bearing into the arm and locking it with a pin
and/or retainer ring. For press-fitting, induction
hardening is applied to ensure hardness for specificparts of the bearing. The FA-treatment technique canbe applied to the latter method.

4.1 External appearance and microstructure of
rocker arm bearing[1] Photos of the rocker arm bearings are shown in
Photo 1.
Each bearing is comprised of an outer ring,

Photo 1 Roller Rocker Bearing

Bearing steel subjected to standard heat treatment

Bearing steel subjected to FA treatment

Photo 2 The prior austenite grain boundaries-67-

NTN TECHNICAL REVIEW No.75（2007）

Cumulative percent failed %5.1 Crankshaft supporting split needle roller bearing
95
90

Being a major component within a main motion
system of any piston engine, the crankshaft converts
the reciprocating motion of the pistons into the rotary
motion via the connecting rod. Conventionally, a split
type sliding bearing is used on the crank journal
section. However, this arrangement leads to a needfor much more oil supply compared with anarrangement that employs a rolling bearing. If a rolling
bearing can be used in the crank journal section, alow-profile structure is necessary. Our crankshaft
supporting split needle roller bearing is comprised of a
two-piece outer ring and a set of needle rollers with a
two-piece cage. The wall thickness of the outer ringcan be as small as 2 mm since we have successfully
developed a unique outer ring splitting scheme that
can help inhibit deformation of the outer ring (Photo
3).

50

3.7 times
10

Conventional
product

FA-treatedproduct

5
10

20

50

100

200

500

1000

Life hContact face pressure GPa

Fig. 7 Comparison of rolling contact fatigue of bearing

2.9

FA-treatedproduct

2.5

13% UP
2.1

Conventional
product
1.7
0

50

100

150

200

250

Life hFig. 8 Comparison of rolling contact fatigue of bearing
by face pressure difference

Photo 3 Split type needle roller bearing for crankshaft

5. Low-friction capable needle roller
bearings (for supporting crankshaft
and camshaft)
There is an increasing need for rolling bearings forsupporting the shafts on automotive engines to reduce
the friction possibly occurring between the shafts.
Since crankshafts and camshafts have offset bearing
journals, it is difficult to install an ordinary non-split
bearing in the axial direction. At the same time,
crankshafts and camshafts require a very low profile
construction. Therefore, a split type sliding bearing
can be used where the housing and shaft slide witheach other.
In order for a crankshaft and a camshaft to readily
incorporate a rolling bearing, we have developed asplit needle roller bearing that provides easy
installation and a low-profile structure. This uniqueneedle roller bearing is described below.

5. 1. 1 Performance evaluation
Performance of our needle roller bearing wasevaluated under the conditions that simulated those ofan actual automotive engine. The torque of our
product was compared with that of a split type sliding
bearing. Also, our product was subjected to a bearing
life test, acoustic measurement, and a measurement
for correlation between oil level and torque. The resultof the torque measurement is described below.

5. 1. 2 Bearing specification
Our needle roller bearing being tested measured 44
mm in inside diameter, 58 mm in outside diameter, 18mm in width, with a roller diameter measuring 3.5 mm,roller length 14.8 mm with 26 rollers. This bearing wascompared with a split type sliding bearing (inside dia.
46 mm, outside dia. 50 mm, width 14 mm).

-68-

Engine Part Technical Trends and New Engine Products

5. 1. 3 Measuring conditions

5. 2 Camshaft supporting split needle roller bearing

The measurement was performed with the NTN
high-speed outer ring run tester. The torquemeasuring conditions for our crankshaft supporting
split needle roller bearing are summarized in Table 3.

The camshaft, rotating in synchronization with thecrankshaft, opens and closes valves by the motions of
the tappets and rocker arms. Conventionally, thejournal section on the camshaft uses a split typesliding bearing; and, generally, when the cylinder
heads are made of aluminum, the camshaft is allowed
to slide relative to the housing by utilizing the bearing
metal function of aluminum itself.
Our camshaft supporting split needle roller bearing
is comprised of a pair of outer rings and needle rollers
with a single-split resin cage. The wall thickness of theouter ring can be as small as 0.6 mm because of apress-formed outer ring. (Photo 4).

5. 1. 4 Measurement result
The running torque measurement results areillustrated in Fig. 9, and the starting torquemeasurement results are shown in Fig. 10. A pictureof the low-friction capable needle roller bearing
evaluation test rig is shown in Fig. 11.
Compared with the split-type sliding bearing, thestarting torque with our needle roller bearing is 90% to
92% lower and the running torque is 39% to 79%
lower.

Table 3 Split type needle roller bearing for camshaft torque measuring situations
Running torqueStarting torqueNTN high-speed outer ring run tester

Test rig

1,500N

9,150N

Applied load

1,000，2,000，3,000，6,000min-1

Speed

Initial acceleration from 0 to 1,000 min-1 in 0.5 sec. period
Engine oil 0W-20

Oil type

Lubricating condition 100˚C, circulating lubrication at rate of 400 mL/min

Ordinary temperature, circulating lubrication at rate of 30 mL/min

10
2
Sliding bearing

1.5

Starting torque N・mRunning torque N・mRolling bearing

1

0.5

0

1000

2000

4000

8

6

4

2

6000

Speed min-1
0

Fig. 9 Result of rotation torque measurement(Split type needle roller bearing for crankshaft)Rollingbearing

Slidingbearing

Fig. 10 Result of start torque measurement(Split type needle roller bearing for crankshaft)

Test bearing
Drive motor

Torque meter

Fig. 11 Over view of test machine for “Low frictionneedle roller bearing”

Photo 4 Split type needle roller bearing for camshaft-69-

NTN TECHNICAL REVIEW No.75（2007）

5. 2. 1 Performance evaluation

5. 2. 4 Measurement result

Performance of our novel needle roller bearing wasevaluated under the conditions that simulated those ofan actual automotive engine. The torque of our
product was compared with that of a sliding bearing ofsimilar size. Also, our product was subjected to abearing life test, acoustic measurement,and ameasurement for mechanical strength of the outer ring
positioning projection. The result of the torquemeasurement is described below.

The running torque measurement result is illustrated
in Fig. 12, and the starting torque measurement resultis shown in Fig. 13.
Compared with the sliding bearing of a similar size,
the starting torque with our needle roller bearing is
92% to 95% lower and the running torque is 35% to
62% lower.

5. 2. 2 Bearing specification

Various development efforts have been made forbetter fuel efficiency with automotive engines, and theexamples include hybrid engines, biofuel engines,
compact powerful engine through downsizing; direct
injection and lean-burn arrangements. To prepare forthe 2015 fuel regulations, further attempts for more
improved fuel economy are expected.
New auxiliary products for automotive engines
featuring new functions will be needed while theconventional technologies are further improved(reduction in pump loss by utilization of valve
continuous variable mechanism; reduction in friction,lighter, more compact arrangements for engine
auxiliaries through use of roller rocker).

6. Conclusion

Our needle roller bearing being subjected to theevaluation measured 23 mm in inside diameter, 31.2
mm in outside diameter, 16 mm in width, with a roller
diameter of 3.5 mm, roller length of 10.8 mm and thenumber of rollers was 16. This bearing was comparedwith a sliding bearing of similar size (inside dia. 23mm, outside dia. 30 mm, width 25.8 mm).

5. 2. 3 Measuring conditions
The measurement was performed with the NTN
high-speed outer ring run tester. The torquemeasuring conditions for our camshaft supporting split
needle roller bearing is summarized in Table 4.

Table 4 Split type needle roller bearing for camshaft

torque measuring situations

Running torqueStarting torqueNTN high-speed outer ring run tester

Test rig
9,150N

Applied load

1,500N

1,000，2,000，3,000，6,000min-1

Speed

Initial acceleration from 0 to 750 min-1 in 0.5 sec. period
Engine oil 0W-20

Oil type

Lubricating condition 100˚C, circulating lubrication at rate of 400 mL/min

Ordinary temperature, circulating lubrication at rate of 200 mL/min

3.5

0.50

3.0

Running torque N・mStarting torque N・mRolling bearing
0.40

Sliding bearing

0.30

0.20

2.5
2.0
1.5
1.0
0.5

0.10

0

Rollingbearing

Slidingbearing

0.00

Applied load→ 1000N

2000N

2000N

2500N

Speed → 750min-1

1500min-1

3000min-1

4250min-1

Fig. 13 Result of start torque measurement(Split type needle roller bearing for camshaft)

Fig. 12 Result of rotation torque measurement(Split type needle roller bearing for camshaft)-70-Engine Part Technical Trends and New Engine Products

To remain compliant with any more demanding
environmental regulation in the future, we willcontinually develop and market our versions of ecofriendly bearing products for automotive engines.

References
1) Website of Ministry of Land, Infrastructure andTransport Japan
2) Industry and Science Systems Co., Ltd., Engine
friction reduction technology text, Engine Friction
Reduction Technology, Nissan Motor Company,
Powertrain Division Headquarters, Yoshinari Fujita
3) Tribology Monthly, Oct. 2003, Special Report II Automobile Tribology: Engine Edition, About the
reduction of engine friction, Suzuki Four-Wheel Drive
Powertrain Design Group, Shun Kosugi, Ken Katagiri
4) Website of Nissan Motor Co., Ltd.
5) Satoshi KITANO, Tadahisa TANAKA, Tomokazu
NAKAGAWA: The Auto-tensioner Market andTechnical Trends, NTN Technical Review No.73(2005)
6) Satoshi KITANO, Tadahisa TANAKA, Tomokazu
NAKAGAWA: The Auto-tensioner Market andTechnical Trends, Tribology Monthly, No.206 (2004)

Photos of authors

Yoshiaki RYOUNO

Kouichi ONIMARU

Kenichi KAWABATA

Shinji OOISHI

Automotive Engineering Dept.
Automotive Sales
Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters

Needle Roller Bearing
Engineering Dept.
Automotive Sales Headquarters

Needle Roller Bearing
Engineering Dept.
Automotive Sales Headquarters-71-

NTN TECHNICAL REVIEW No.75（2007）[ New Product ]Ball-screw Unit for Variable Valve Event and Lift System

Keisuke KAZUNO*

Demand for improved fuel economy and cleaner exhaustemissions is increasing. One well-known technology to meet thisdemand is VTC, which makes it possible to change the timing
between the crankshaft and engine valve opening. Recently,
demand to continuously control the amount of valve lift has grown inorder to further protect the environment. NTN has developed a newball-screw unit, which reduces friction significantly while still having
good durability. This ball-screw unit can control the electric actuatorwith higher efficiency and reaction time than current systems.
This article introduces the features of this ball-screw unit.

achieves an overall smaller size, higher reliability(safety) and quicker response that conventional
systems. Working together with Hitachi, Ltd., NTN has
developed a unique ball-screw unit that operates thedrive mechanism for the actuator that is responsible
for the continuous variable control.
This paper hereunder describes this ball-screw unit.

1. Introduction
Currently, most automotive gasoline engines are
four-stroke cycle engines in which one cycle consistsof four steps: induction of air-gasoline mixture into theengine, compression of the mixture, ignition of thecompressed mixture, and emission of the resultant
exhaust. Thus, this is how the gasoline engine
develops power.
The valves situated at the top of the engine govern
this operation. By varying the timing for opening andclosing these valves, or by changing the lift in verticalmotion of the valves, the fuel economy or output of thevehicle can be improved.
Recently, there has been mounting pressure for
reduction in CO2 emissions. Development work has
been progressing for a continuously variable valve
mechanism that is capable of the more delicate
control that is necessary with the operating conditionsof cars. In Europe, a worm-gear-based continuously
variable valve mechanism has already been in themarket for several years. Nissan Motor Co., Ltd. andHitachi, Ltd., have succeeded in jointly developing the
Variable Valve Event & Lift (VEL) system, which helps
reduce electricity consumption and produces a quicker
response. Utilizing the fact that a ball-screw is low
friction and is highly efficient, this mechanism

2. Structure of VEL
2. 1 Appearance of VEL
An illustration of the VEL is shown in Fig. 1.
Rotary motion of the driveshaft, which includes an
eccentric cam, is translated into the vertical motion of
the valve lifters by a plurality of links (link A—rocker
arm—link B—output cam).
Based on this principle and by varying the phase of
the control shaft coupled with the actuator section, thelift and operating angle of the valves are invariably
controlled.
Incorporation of the VEL system will lead to
improved output and response of automotive engines,
improved fuel economy and cleaner emissions. In
particular, the fuel economy appears to improve byapproximately 10%.

*Automotive Engineering Dept. Automotive Sales Headquarters-72-Ball-screw Unit for Variable Valve Event and Lift System

2. 2 Structure of actuator for VEL
The structure of the actuator section is
illustrated in Fig. 2.
A ball-screw unit, as shown in Photo 1,
is used on the actuator section in order toachieve a quicker response and lower
electricity consumption. A DC motor islinked to the ball-screw shaft with acoupling. The rotating power of the motor isefficiently converted into the linear motionof the nut owing to a smaller rolling
resistance, which is an advantage of theball-screw unit, and accordingly varies theangular phase of the arm via the linkconnected to the middle point of the nut.
Being engaged with the control shaft end,
this arm invariably controls the lift andoperating angle of the valves.
Incidentally, it is important to inhibit thenoise and vibration that result from vibrationloads occurring from the movement of thevalves. To address this issue a highefficiency ball-screw is used, which features
an optimized clearance. Also, thesupporting bearing used is a four-point
contact ball bearing that is capable of acompact, lightweight arrangement with asmaller clearance.

3. Features of ball-screw unit
for VEL

Actuator

sectionBall-screw unit

Rocker arm
Control shaft
Link A

Output cam

Driveshaft

Link B
Valve lifter

Valve※ Courtesy Hitachi, Ltd.

Fig. 1 Appearance of VEL system

4-point contact ball bearing
Ball-screw nut
Ball-screw shaft
Deep grooveball bearing

DC motor

Housing Link

An illustration of the ball-screw unit for
the VEL system is shown in Photo 1.

Arm※ Courtesy Hitachi, Ltd.

Fig. 2 Structure of actuator

4-point contact ball bearing

Disk deflector Nut

Photo 1 Ball screw unit for VEL system-73-

Ball-screw shaft

NTN TECHNICAL REVIEW No.75（2007）

The ball-screw unit on the VEL system canpositively withstand the load application from thevalves and still features a lower cost, as the result ofincorporating the following unique features:(1) Use of disk deflector(2) Integration of supporting bearing by staking(3) Adoption of oil-immersed induction hardening
technique for ball-screw shaft(4) Formation of nut thread groove without using
grinding techniqueFurther detail on these features is given in thefollowing sections:Compared with the conventional oval deflector, theextraction load for the disk deflector is approximately
1.8 times as great.

Wing

3. 1 Use of disk deflector
The ball-screw unit uses a deflector system forcirculating the balls, wherein the deflector used is adisk type (improved shape) instead of a conventional
oval type. An illustration of a conventional oval
deflector is given in Photo 2, while the new disk typedeflector is shown in Photos 3 and 4.
The new deflector is inserted into the round holeformed on the nut through the bore of nut, wherein thewing sections are allowed to remain seated in thethread groove where the balls do not roll, and several
areas on the outer circumference of the nut are
secured by staking.
Because the previous deflector was an oval type, it
was necessary to open the deflector hole on the nut
by moving an end-milling cutter sideways from theouter circumferencial side of the nut in order to obtain
the oval hole. In contrast, with our ball-screw unit, thedeflector hole is round-shaped and can be easily
machined with a standard tool; thereby the deflector
hole is made at a lower machining cost.
Furthermore, the new deflector is made of amaterial boasting a sufficient mechanical strength sothat the force needed to extract the deflector from thenut is sufficiently great relative to the load beingexerted from the valve.
The result of the comparison between theconventional oval deflector and the new disk deflector,
in terms of the force needed for extracting thedeflector from the nut, is plotted in Fig. 3.

Photo 3 Appearance of new deflector

Staked areas

Photo 4 Appearance of new deflector staking

2.5

2Load kN

Disk deflector
1.5

1

0.5

Oval deflector
0

Oval (conventional) deflector

0

0.1

0.2

0.3

0.4

Deflection mm

Fig. 3 Load-Displacement curve of deflector

Photo 2 Appearance of past deflector

-74-

0.5

Ball-screw Unit for Variable Valve Event and Lift Systemapproximately 10 times as great as the maximum axial
load acting on the ball-screw shaft.
Also, an extraction fatigue test was performed
against the vibration load acting on the staked areas
on the bearing.
The vibration load specified in Table 1 was applied
to the ball-screw shaft, thereby verifying that thebearing did not come off the ball-screw shaft.

3. 2 Integration of supporting bearing by staking
When being fixed along the axial direction on a ballscrew shaft, a bearing is usually secured with a locknut or with a retainer ring. The locking arrangementwith a lock nut means a necessity of an additional part(lock nut) and a procedure for forming the thread forthe nut on the ball-screw shaft, which increases cost.
This arrangement further needs an anti-loosening
feature for the lock nut. Incidentally, a countermeasureagainst wear and tear will be needed for the retaining
arrangement when using a retainer ring if a vibratingload acts on the retainer ring.
For the new ball-screw unit, the bearing is fixed tothe shaft through plastic-deformation of a relevant
portion of the ball-screw shaft by staking. Illustrationsof the staked area are shown in Photos 5 and 6.
As shown in Photos 5, a portion of the shaft overwhich the bearing had been inserted was plasticdeformed so that it was brought into contact with theend face of the bearing, thereby integrating thebearing with the ball-screw shaft.
However, if the bearing is more strongly staked toensure sufficient fastening strength, the circularity ofthe bearing inner ring could be adversely affected. To
address this issue a spacer is inserted between thestaked area and the bearing end face to decrease theload possibly occurring from radial deformation, which
results from staking. Thereby, deformation of the
bearing inner ring is avoided.
The result of the evaluation for the axial extraction
force needed to remove the bearing secured by thisstaking method is plotted in Fig. 4. This evaluation
was performed by applying a load to the bearing end
face while the shaft was locked and then measuring
the bearing displacement. The bearing extraction
force needed was approximately 7 kN, which is

12

Load kN

10

8

6
4
2
0
0

0.1

0.2

0.3

0.4

Displacement on bearing mm

Fig. 4 Load-Displacement curve of bearing

Table 1 Table of fatigue test condition of bearing coming out
Load
(kN)

Frequency(Hz)0.79±0.77
1.57±1.5530

2.35±2.33
3.13±3.11

20

Test duration Number of loading cycles(ｈ)(×107)
103.5

1.11

24

0.25

24

0.25

36

0.25

SpacerStaked areas

Photo 5 Appearance of bearing staking（view from axial side）

Photo 6 Appearance of bearing staking（view from radial side）-75-

NTN TECHNICAL REVIEW No.75（2007）

NTN has realized this achievement throughadoption of a NC-based, high-precision cutting
process.
After the high-precision cutting work the nut issubjected to carburizing and quenching. As a result of
this process, a grain boundary oxidized layer occurs
on the thread groove surface owing to the oxygen in
the furnace. If the groove is used as the ball-screw
groove without any modification, then the thread
groove will pose problems in terms of wear resistanceand durability.
To overcome these problems the nut-rolling surfaceof the ball-screw unit, after undergoing the carburizing
and quenching processes, is treated with a special
shot-peening process to remove the boundary grainoxidized layer. This improves the wear resistance anddurability of the thread groove.
NTN has performed a retaining & durability test inorder to evaluate the wear resistance of the new ballscrew unit. In this test, we applied a vibrating load tothe ball-screw shaft whose position was retained while
the shaft was not allowed to rotate. Then the increasein the clearance on the ball-screw was evaluatedrelative to the clearance before the test. Table 2
summarizes the comparison result in terms of wear
resistance between samples that have undergone a
special shot-peening process and those that have notundergone this process.

3. 3 Adoption of oil-immersed induction
hardening technique for ball-screw shaft
Previously, the ball-screw shaft was inductionhardened in air, which resulted in oxidation scaleoccurring on the rolling surface of the thread and soan additional process was necessary to remove thisscale.
In the induction hardening process for the ballscrew shaft of the NTN ball-screw unit, the newly
developed ball-screw shaft and induction-hardening
coil are immersed in oil to isolate the rolling surface
from the atmosphere. Photo 7 shows the state ofoxidation scale deposition on the ball-screw shaft
subjected to conventional induction hardening in
atmosphere and on the newly developed ball-screw
shaft that has undergone induction hardening in oil.
Virtually no oxidation scale will occur from the oilimmersed induction hardening process and so anoxidation scale removal process can be deleted. Thus,
the so-obtained ball-screw shaft boasts excellent
surface quality and the ball-screw unit features a lower
price.

Shaft subjected
to conventional
induction
hardening

Shaft subjected
to oil-immersed
induction
hardening

Photo 7 Comparison of Oxidation scale on the shaft

3. 4 Formation of nut thread groove without
using grinding techniquePreviously, the machining process for the thread
groove on the ball-screw nut consisted of a roughformation of the thread groove while leaving asufficient grinding allowance. This was followed by aheat treatment process and then a grinding process toform the final thread groove profile.
Though positively capable of attaining an accurate
thread profile, the thread grinding processnecessitates an expensive plant and a longer
processing time, leading to a cost issue.
The new ball-screw unit employs a novel thread
groove machining process for the nut that offers
sufficient durability and operability of the nut without
involving grinding work. Photo 8 shows the rollingsurface of the nut machined without grinding.

Photo 8 Surface of no-grinding nut groove

Table 2 Comparison of wear resistance
Test ball-screwSpecimen
undergone
special shot peening

Test load

Radial and axial loads combined
1032N，40Hz

Lubrication

Engine oil (5W-30)）
1cc/5 min., drip

Test duration (h)68Increase in clearance
on ball-screw (mm)-76-

Specimen notundergone
shot-peening

0.008

0.001

Ball-screw Unit for Variable Valve Event and Lift System

Cumulative percent failed %Samples that have undergone the special shotpeening process exhibit virtually no increase inclearance and have good wear resistance.
Next, we compared the durability of the ball-screwthat has undergone the special shot-peening processwith that of the ball-screw has not undergone a similar
process; the comparison result is graphically plotted inFig. 5. In the durability test a constant load of 2400 N
was applied to each specimen until flaking occurred
and then a comparison was made of the life results.
The L10 life of the ball-screws that did not undergo
the special shot-peening process stood at 9.4x106
rev., while the specimens that had undergone thisprocess was 15.4x106rev. This means that the life of
the samples that had undergone the special shotpeening process are approximately 1.6 times as longcompared with those that had not undergone thisprocess.
On a bench test with an actual automotive engine,
the new NTN ball-screw unit successfully withstood adurability test equivalent to 300,000 km of actual
vehicle travel. This means that the ball-screw unit
satisfies the requirements for both a high degree of
durability and a lower cost.

4. Conclusion
The new NTN ball-screw unit incorporates various
novel design techniques and machining processes
has been described above.
Efforts have been mounting for electromechanicalarrangements not only for variable valve mechanisms
whose market seems to expand dramatically, but also
for various car components. Under this situation more
car developers are thinking of ball-screw based drive
mechanisms. Therefore, the market size for improvedball-screw mechanisms for automobiles will continue
to expand.
NTN has been promoting new developmental workfor the ball-screw units that achieve a lower cost andimproved durability and will remain committed to
expanding the applicability of ball-screw units forautomobiles.
In closing this document, I want to express mygratitude to the people of Automotive System Group,
Hitachi, Ltd., who have kindly supported thedevelopment work for our ball-screw products andprovided their invaluable engineering information.

Specimen not undergone
shot-peening

95
90
80
70
60
50
40
30

Specimen undergone
special shot peening

20
10
5

4

5 6 7 8 9 1
106
107

1.5

2

3

4

5

Total rotations

Fig. 5 Comparison of durability

Photo of author

Keisuke KAZUNO
Automotive Engineering Dept.
Automotive Sales Headquarters

-77-

6

NTN TECHNICAL REVIEW No.75（2007）[ Technical Paper ]Development of
an End-Pivot Type Mechanical Lash Adjuster

Eiji MAENO*
Hiroshi BUNKO**
Katsuhisa YAMAGUCHI**NTN has developed a Mechanical Lash Adjuster (MLA) that can replace
conventional Hydraulic Lash Adjusters (HLA). This MLA, which applies
buttress threads as the adjustment mechanism, shows excellent
performance in fundamental function tests on valve-lift stability, valve-lashadjustment speed, and low-temperature characteristics. Durability tests
confirmed that the MLA is capable of a 20% shorter axially design incomparison with conventional HLAs.

1. Foreword

2. Structure

Recently, an increasing number of car models areincorporating roller rocker arm type valve systems with
one of the goals being better fuel economy. However,compared with a direct type valve system a rocker arm
type valve system has a more complicated structure
and a greater number of wear points. Therefore, arocker arm type valve system needs a mechanism foradjusting the valve clearance 1).
Common examples of valve clearance adjustment
mechanisms are a fixed type and a HLA. The fixed
type valve clearance adjustment mechanism is
manually adjusted with a screw and then tightened tolock the adjustment, where as the HLA is an automatic
adjustment system. However, these conventional
valve clearance adjustment mechanisms havedrawbacks. The fixed type needs to be adjusted
manually when necessary. The drawback with the
HLA is that when bubbles in the engine oil enter thehigh-pressure chamber a decrease in rigidity andnoise can occur 2). NTN has developed a uniquemechanical lash adjuster (MLA) that supersedes theseconventional mechanisms. This report describes thebasic MLA functions and the result of their durability
testing.

Fig. 1 shows a cross-sectional view of the MLA andFig. 2 gives a schematic view of the MLA incorporated
into an automotive engine. Using buttress thread
screw technology in conjunction with a return spring,the MLA is able to provide automatic adjustment of thevalve clearance.
On the buttress thread screw system the malethread is formed beneath the pivot member thatserves as the fulcrum for the rocker arm, while thefemale thread is in the inner circumference on thecase that houses the pivot member. By operation of acompressed coil spring the pivot member is forced
upward toward the rocker arm side, while a springseat, whose top end is spherical, is inserted between
the pivot member and the compressed coil spring todecrease the running resistance on the MLA.

**Automotive Engineering Dept. Automotive Sales Headquarters**New Product Development Dept. New Product Development R&D Center-78-

Development of an End-Pivot Type Mechanical Lash Adjuster

axial load acting on the buttress thread screw. In otherwords, by appropriately adjusting the lead angle andtwo flank angles, it is possible to allow the buttressthread screw to remain at a self-sustained stateagainst an axial load working from one side and toslide and rotate by an axial load acting from the
opposite side. Utilizing this direction-dependency of abuttress thread screw, the NTN MLA realizes theautomatic valve clearance adjustment function.
The shape of the buttress thread screw is illustrated
in Fig. 3. The angles of the thread ridge are designed
such that the screw slides in the projection direction
for the pivot, and remains at a self-sustained state in
the drive-in direction for the pivot, and that the screw
does not develop wedge engagement-induced seizure
against a drive-in load. The conditions required for the
thread angles are summarized with the formula (1)below:
tanα＞μmax・secθ2' (projecting condition)tanα＞μmin･secθ1' (self-sustaining condition) ………(1)cotθ1＞μmax・cosα (wedge angle)
where,
α: lead angle
θ1’ : flank angle (self-sustaining side) on a crosssectional plane A that is square to the threadlead
θ2’ : flank angle (sliding side) on a cross-sectional
plane A that is square to the thread lead
μmax : maximum friction coefficient for friction
between thread faces
μmin : minimum friction coefficient for friction
between thread faces

Pivot

Case

Spring seat

Return spring

Cap

Fig. 1 Cross-section of MLA

CamRocker arm｝

MLA
Engine valve

Fig. 2 Installation layout of MLA

Formula (2) shows the relationship between theflank angle q’ on the cross-sectional plane A square to
the thread lead and the flank angle q on the crosssectional plane B that includes the screw axis.
However, note that the buttress thread screw used on
the MLA features a sufficiently small a that permits anapproximation of cosα≒1.

3. Operating principle and functions
3. 1 Operating principle (about buttress thread
screw)The static self-sustained (no occurrence of sliding
rotation against the axial static load applied) conditionof a screw is governed by the relation between theangles of the thread ridge (lead angleαand flank angle
θ'on a cross-sectional plane square to the threadlead) and the inner-thread face friction coefficient.
More specifically, if the actual friction coefficientμ is
greater than the static self-sustaining friction
coefficient μs defined by tanαcosθ', then the screw
will remain at a self-sustained state; and when theformer is smaller than the latter, the screw will slide 3).
Because the two neighboring thread faces on abuttress thread are not symmetrical with each other,
the self-sustaining coefficient ms of the buttressthread can vary depending on the orientation of thetanθ1＝tanθcosα ………………………………(2)
A

B

θ1
θ2
α
Fig. 3 Shape of the buttress threads-79-

NTN TECHNICAL REVIEW No.75（2007）

3. 2 Functions

3. 2. 2 Adjustment against enlarged valve clearance

In Fig. 4, typical factors that affect the valve
clearance are illustrated. The lash adjuster is required
to expand and/or contract in accordance with thevariation in the valve clearance, which results fromthese factors, to maintain the valve clearance to alimited range.

While the valve clearance increases due to factors
including thermal expansion on the engine head andwhen the valve opens, the motion of the buttressthread screw is identical to that in the steady operation
mode described in Sec. 3.2.1. On the other hand,when the valve closes, the pivot rotates in the return
direction by a certain angle (an angle for slight sliding
between the thread faces when the valve has opened
plus an angle equivalent to a newly enlarged valve
clearance), thereby the MLA expands causing thevalve clearance to be maintained at an adequate level(Fig. 6).

3. 2. 3 Adjustment against decreased valve clearance
A worn valve seat will lead to reduced valve
clearance. Since the valve clearance further
decreases in such a situation due to the backlash on
the screw; consequently, when the valve closes, thepivot cannot move upward to a position where thesliding side thread faces come into contact with eachother and no return rotation occurs on the pivot.

Thermal expansion
on valve

Thermal expansion
on engine head
Wear on valve seat

Axial displacement of pivot

Rotation of pivot

Fig. 4 Factor of changing valve clearance
Cam load

Amount of returnrotation increases.

Self-sustaining side threadface comes into contact.

Sliding side thread facecomes into contact.

（a）When valve opens（b）When valve closes

3. 2. 1 Steady operation mode
When the valve opens as the thread faces on the
self-sustaining side come into contact with each otherdue to the input from the cam load, which results inrelative sliding occurs between the thread faces
because of a squeeze film and elastic deformation of
the thread faces 4). When the valve closes, the pivot islifted by the reaction force of the return spring until thesliding side thread faces come into contact with eachother, thereby maintaining the MLA at a constantheight (Fig. 5).

Axial displacement of pivot
Cam load

Pivot(male thread)Fig. 6 Motion of buttress-thread (expansion)

Rotation of pivot

Axial displacement of pivot

Amount of return rotation= amount of sliding when valve opens

Cam load

Rotation of pivot

No return rotation occurs.

Case(female thread)

Self-sustaining side threadface comes into contact.

Sliding side thread facecomes into contact.

（a）When valve opens（b）When valve closes

Self-sustaining side threadface comes into contact.

（a）When valve opens

Fig. 5 Motion of buttress-thread (steady)Sliding side thread facecomes into contact.

（b）When valve closes

Fig. 7 Motion of buttress-thread (contraction)-80-

Development of an End-Pivot Type Mechanical Lash Adjuster

Because of this, the slight sliding occurring when thevalve opens will accumulate, thereby the MLA will
gradually contract and as a result, the valve clearance
will be maintained at an appropriate level (Fig. 7).

was used to measure the area around the valve lift
peak and the other for measuring the area around thevalve seat. The sensor configuration is illustrated inFig. 9 and the test conditions are summarized inTable 1.

4. Evaluation result

4. 1. 2 Result

4. 1 Valve-lift stability
4. 1. 1 Analysis method

Valve lift peak mm

To investigate the valve-lift stability in steady
operation mode, variation in the valve lift peak wasmeasured while the engine speed was maintained at aconstant level and by sweeping the engine speed(repeated variation). A picture of the engine bench test
rig is shown in Fig. 8. The crankshaft of an automotiveengine (1500 cc) was rotated with an electric motor
and the valve position was detected with eddy current
type gap sensors. Since the detection range of thecompact gap sensor is limited two sensors were used
to measure variation in the valve lift peak. One sensor

9.6

180

9.8

160

10.0

140

10.2

120

10.4

100

10.6

80

MLA (Cylinder_1)

10.8

HLA (Cylinder_3)

60

11.0

Oil temperature

40

11.2
0

2000

Fig. 8 Engine bench for functional tests

4000

Temperature ˚C

The valve lift peak positions are continuously plottedin Figs. 10 and 11. Fig. 10 represents a plotting
obtained from an operation where the engine was run
at a constant speed, while for the plot in Fig. 11, theengine speed was swept. For the test, the MLA(NTN’s new product) and the HLA (conventional
product) were both commonly incorporated into one
automotive engine, thereby they were simultaneously
analyzed. The absolute values of valve lift height aresmaller with the MLA owing to the backlash on thescrew in the MLA. However, the magnitude ofvariation in valve lift peak is the same with both the
HLA and MLA. Thus, it was verified that the stability
with a MLA is sufficiently high.

20
6000

Time sec

Fig. 10 Valve lift peak (constant-rotation)

Valve lift peak position
measuring gap sensor

9.6

16

9.8

14

10.0

12

10.2

10

10.4

8

10.6
10.8
11.0

Fig. 9 Layout of gap-sensors

11.2
0Table 1 Condition of valve-lift stability test
Oil viscosity

Run duration

2000

4000

4
2
0
6000

Time sec

SAE OW-20

Fig. 11 Valve lift peak (cyclic-rotation)

Oil temperature Approx. 40-130˚C (natural temperature increase)
Crank speed

6

MLA (Cylinder_1)
HLA (Cylinder_3)
Crank rotation

1 6000 min-1 constant
2 600-6000 min-1 sweeping
6000 sec-81-

Rotation ×1000 min-1

Valve lift peak mm

Valve seating position
measuring gap sensor

NTN TECHNICAL REVIEW No.75（2007）

seat (as shown in Fig. 13) and then by analyzing theresultant valve behavior (seating position and peak
position of the valve).

4. 2 Expansion/contraction follow-up quality
4. 2. 1 Analysis method
After the engine is started the exhaust valve, being
exposed to hot exhaust gas, undergoes thermal
expansion. When the exhaust valve expands, thevalve clearance will decrease. In the case of a HLA,
the original valve clearance on it is 0. If contraction of
the lash adjuster is too slow the HLA can develop
compression leakage. With the MLA the backlash on
the buttress thread screw functions as a valve
clearance, thereby the valve expansion iscompensated for and a compression leak does notoccur immediately. However, expansion-contraction
follow-up quality is an important characteristic for the
performance of the lash adjuster. The expansioncontraction follow-up performance of the MLA wascompared with that of the HLA. An example of the test
bench used is shown in Fig. 12. The engine head wasplaced on the test bench and the camshaft was driven
by an electric motor via a timing belt. The position of
the valve lift was detected from beneath the valve witha laser displacement meter. The test conditions are
summarized in Table 2.
The displacement on the lash adjuster due toexpansion/contraction was measured by inserting andremoving a shim between the valve face and valve

4. 2. 2 Result
Valve behaviors in terms of the seating position andpeak position, with a 0.3 mm shim inserted or
removed, are illustrated in Figs. 14 and 15.

Sliding on self-sustainedside thread faces (backlash = 0)

Valve lift peak mm

10.8
10.7

Restoration of backlash

10.6
10.5
10.4

MLA elongates, causing
the lift peak to be higher.

10.3
10.2
10.1
0

5

10

15

20

25

Time sec（a）Peak position

30

Valve lift peak mm

0.5

Shift of seating
position by the
thickness of shim

0.4
0.3
0.2

Variation by backlash

0.1
0
-0.1

Shim inserted Shim removed
0

5

10

Valve fully CLOSED
15

20

25

30

Time sec（b）Seating position

Fig. 14 Valve behavior in adjustment test（MLA）
11

Valve lift peak mm

Laser displacement
meter

Fig. 12 Test bench for functional test
Table 2 Condition of adjustment speed test

10.8
10.7
10.6
10.5

Lift peak will become
higher owing to
expansion on HLA.

10.4
10.3

Oil viscosity

SAE OW-20

Oil temperature

Approx. 40/80°C (temperature controller)

Crank speed

600，1000，2000 min-1

Run duration

30 sec

0

5

10

15

20

25

30

Time sec（a）Peak position
Valve lift peak mm

0.5

Engine valve

Shim

Leak-down

10.9

Valve seat
Manifold

Shift of seating
position by the
thickness of shim

0.4
0.3
0.2
0.1
0
-0.1

Shim inserted
0

Shim removed
5

10

Valve fully CLOSED
15

20

25

30

Time sec（b）Seating position

Valve faceFig. 15 Valve behavior in adjustment test（HLA）

Fig. 13 Thim insertion diagram-82-

Development of an End-Pivot Type Mechanical Lash Adjuster

For both the HLA and MLA with the shim inserted,
the peak position varies at almost the same time as
when the valve seating position varies. This meansthat an increase in valve clearance resulting from theinserted shim is compensated for as the lash adjuster
instantaneously expands. In other words, theexpansion speed of the lash adjuster is sufficiently
high relative to the thermal expansion and wear
speed.
After the shim is removed, the HLA graduallycontracts and the seating and peak positions return totheir original positions at a same speed. When theshim is removed with the MLA, the screw backlash iseliminated and the pivot is gradually driven in. Then,
when the valve can be fully closed the pivot is further
driven in by a distance equivalent to the screw
backlash; thereby the MLA resumes its original state.
The contraction speed is relatively low and this factsometimes leads to an occurrence of compression
leakage. The HLA was compared with the MLA in
terms of the time needed for the lash adjuster to sinkby a particular stroke (contraction time) and the resultis shown in Fig. 16. Since the HLA contracts due tothe oil leakage from the high-pressure chamber, its
contraction time varies depending on the temperature.
In contrast, the contraction time of the MLA isdependent on the running speed rather than thetemperature. This is because the contraction on the
MLA derives from accumulation of slight sliding on thethread faces occurring in each cycle. The follow-up
speed of the MLA is slowest in the idling mode.
However, this slowest speed is still sufficiently highrelative to the thermal expansion speed of the valve.

4. 3 Low temperature starting quality
At lower temperatures the viscosity of oil getshigher. Due to lower temperatures the female threadface is closely situated to the male thread face, which
results in an oil film more readily occurring between
these thread faces (squeeze effect). If the fluid
lubrication state is present between the thread faces
due to this oil film (squeeze film), the friction
coefficient will greatly decrease and valve lift loss toover-rotation of the screw can occur. With the MLA the
thread faces are provided with both oil drain grooves
and fine concavities/convexities in order to be able todissipate the squeeze film between the thread faces ina short time and attain a mixed lubrication mode
where a relatively large frictional force is obtained.
Valve lift behavior in low temperature starting operation
was analyzed for samples that each featured a uniquethread face specification. The analysis result ispresented below.

4. 3. 1 Analysis method
The test bench for the engine head was cooled to a
constant temperature of -30°C in a cryo-refrigerating
chamber and the displacement in valve lift wasmeasured. The two test pieces used (Samples A andB) each featured a unique thread face roughness andgroove structure. The test conditions are summarizedin Table 3.
Table 3 Condition of low temp. starting test
Oil viscosity

SAE 10W-30

Oil temperature-30˚C (constant temperature chamber)

Crank speed

1500 min-1

Run duration

60 sec4. 3. 2 Result
Adjustment time sec/mm

60

HLA (40˚C)
MLA (40˚C)

50Variations in the peak positions in the valve lift
curves are plotted in Figs. 17 and 18. With Sample A
the valve lift loss of approximately 1 mm occurs
immediately after the engine is started. In contrast, thevalve lift loss with Sample B is 0.1 mm or smaller. The
valve lift curve of seating position obtained from
Sample A is given in Fig. 19, while the seating
position obtained from Sample B is illustrated in Fig.
20. In Sample A the ramp section formed on the highpoint of the cam is not apparent in the valve lift curve;
therefore, it should be understood that the MLA sinks
due to excessive sliding of the screw and the valve
clearance increases to a level higher than the ramp
height. In contrast, with Sample B the evidence of thepresence of the ramp section is apparent, whichindicates that the amount of sliding on the screw isvery small.

HLA (80˚C)
MLA (80˚C)

40
30
20
10

idling speed

0
0

500

1000

1500

2000

2500

-1

Crank rotation min

Fig. 16 Contraction speed of HLA and MLA-83-

NTN TECHNICAL REVIEW No.75（2007）

9.0
9.5

10

Valve lift

10.0

Crank rotation

5

10.5
11.0
10

20

30

40

50

0
60

15
9.0

Valve lift

9.5
10.0

5

11.0
0

10

20

30

40

50

Time sec

Time sec

Fig. 17 Valve lift peak (Sample A)

Fig. 18 Valve lift peak (Sample B)

0.8

0.8

0.7

0.7

0.6

0.6

0.5
0.4
0.3
0.2

0
60

Ramp height

0.5
0.4
0.3
0.2
0.1

0.1
0
60.00

10

Crank rotation

10.5

Valve lift mm

Valve lift mm

08.5

Crank rotation

15

Valve lift mm

Valve lift mm

Valve lift loss

Crank rotation ×1000 min-1

8.5

20

8.0

×1000 min-1

20

8.0

60.05

60.10

0
30.00

60.15

30.05

30.10

30.15

Time sec

Time sec

Fig. 19 Valve lift curve (Sample A)

Fig. 20 Valve lift curve (Sample B)

In conclusion, by adequately designing the surface
roughness of the thread faces and groove structure it
is possible to inhibit excessive sliding between thread
faces even at a very low temperature (–30°C) andattain a stable valve lift.

excessive sliding of the screw did not occur in the lowtemperature starting test. In addition, by varying thearea of thread faces the flank pressures were sorted
into several groups. The test conditions applied aresummarized in Table 4.

4. 4 Flank pressure dependency of wear on
the thread faces

Table 4 Condition of the endurance test
Oil viscosity

With the MLA the male and female thread faces arein contact with each other in a boundary lubrication or
mixed lubrication state; therefore reducing wear on thethread faces poses an engineering challenge.
In an attempt to address this issue an automotiveengine was subjected to a durability test in order todetermine the allowance for thread face flankpressure. From the durability test the wear on thefemale thread face of the case member wasmeasured.

SAE 0W-20

Oil temperature Approx. 110°C (natural temperature increase)
Crank speed

Engine maximum output speed

4. 4. 2 Result
The interrelation between the maximum flankpressure on the thread faces and the wear on thefemale thread face on the case member is illustrated
in Fig. 21. Each maximum thread face flank pressure
was calculated based on the maximum value of theload being an input to the MLA within one rotation of
the camshaft. According to this result, it is possible todetermine the minimum necessary mesh height with

4. 4. 1 Test method
As shown in Fig. 8, the crankshaft of an automotiveengine was driven with an electric motor. The thread
faces of each specimen were preconditioned such that-84-

Wear of thread faceof female screw

Development of an End-Pivot Type Mechanical Lash Adjuster

sufficiently high stability.
2) The contraction speed of the MLA is governed by
the engine speed and is least dependent on thetemperature. The contraction speed is lowest whenthe engine is idling; nevertheless, the lowest
contraction speed is still sufficiently high comparedwith the thermal expansion speed of the valve.
3) It has been verified that by adequately designing
the quality of thread faces the valve lift loss at -30°C
is limited to 0.1 mm or smaller.
4) Based on the relation between the wear on thethread faces and the flank pressure on thread faces,
the minimum necessary mesh height for the
threaded faces is determined. With the mesh height
known it is now possible to optimally design the
MLA that best suits the intended automotive engine.
For example, if our MLA is adopted for a 1500 ccclass engine the axial length needed isapproximately 20% shorter compared with the HLAof the same outside diameter.

Criteria

Upper limit of standard

Maximum pressure of thread faceFig. 21 Wear on the threaded flank of female screw withrespect to maximum flank pressure

the threaded section and obtain a MLA design that is
optimized for an automotive engine application. If
NTN’s MLA is applied to an automotive engine, forexample a 1500 cc class engine, the axial dimensionof the resultant package can be approximately 20%
smaller compared with a package comprising of an
HLA with the same outside diameter as our MLA.

References

5. Conclusion

1) Kimihiko Todo, Yuji Yoshihara: Gasoline Engine:
Roller Arm, Journal of Society of Automotive
Engineers of Japan, No.59 p.29 (2005-2)
2) Noriyuki Miyamura, Syuji Nagano: Hydraulic Lash
Adjuster, Journal of Society of Automotive Engineersof Japan, No.38 p.1110 (1984-9)
3) Akira Yamamoto: Principles and Design of Screw
Fastening, Yokendo, p.30 (1970)
4) Bunko et al.: Fundamental analysis of the dynamic
behavior of buttress-threaded screws, Japan Societyof Mechanical Engineers, 2005 Annual Conference
Proceedings (4), p.147 (2005)

We have evaluated the basic functions anddurability of an automatic valve clearance adjuster(mechanical lash adjuster/MLA) that incorporates abuttress thread screw and obtained the following
findings:
1) When the engine is running at a constant speed(6000 min-1) or when the engine speed is swept(600-6000 min-1), the variation in valve lift peak with
the MLA is comparable with that of the currently
mass-produced hydraulic lash adjuster (HLA). Thus,
we have verified that the MLA has attained

Photos of authors

Eiji MAENO

Hiroshi BUNKO

Katsuhisa YAMAGUCHI

Automotive Engineering Dept.
Automotive Sales
Headquarters

New Product Development Dept.
New Product Development
R&D Center

New Product Development Dept.
New Product Development
R&D Center-85-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Transmission Technology Trends and Product Developments

Takahiro KANAMOTO*
Akihiko KATAYAMA**

Takashi UENO*
Masanori SATOU**

The transmission is an important functional component thattransmits the rotation power generated by the engine to the drive
shaft and the drive wheel. Therefore, high reliability is demandedfrom the transmission bearings.
In addition, the market demands of --low fuel cost (low friction),
reduced size, weight, and long life --have also increased in recent
years along with concern for the environment. For this reason NTN
introduces product developments that satisfy the market demands.

order to achieve low friction, and compact lightweighttransmission designs to help decrease the vehicle’s
weight; in particular, a shorter transmissions so that atransmission can fit in a more compact engine
compartment.
Thus, the requirements that newly developed
automotive transmission products need to satisfy
include reduced running friction (lower torque), morecompact size, lighter weight and resultant longer life.
This paper reports newly developed transmissionrelated products that will satisfy these requirements.

1. Introduction
Automotive transmissions have been generally
categorized into manual transmissions (MT) andautomatic transmissions (AT). However, continuously
variable transmissions (CVT) are also used.
Though the market in Japan for MT seems to beshrinking, MT will remain favored for a while in the EU
market owing to the better fuel economy associatedwith them. AT are most commonly used in theJapanese markets as well as other regions. Recently,
6- to 8- speed AT designs are becoming morecommon on large-sized vehicles because of theircapabilities for better riding comfort and improved fueleconomy. These AT systems of increased speed
ranges will constitute a mainstream of the AT
products. The most outstanding advantage of the CVT
is fuel economy associated with CVT-equipped
vehicles running within the city. Boasting better torquetransmission efficiency, an increasing number of CVT
products are used on compact and middle-sized
vehicles. Recently, the DCT (Dual Clutch
Transmission), which may be regarded as the third
generation automatic transmission system following
the AT and CVT, has been adopted by European car
manufacturers.
The current trends common to these transmission
types include use of low-viscosity transmission oil in

2. FA (Fine Austenite Strengthening)technique
Ingress of foreign matters into a bearing in an
automotive transmission is unavoidable, and bearings
can develop premature flaking that starts at a foreign
matter- derived dent.
By paying attention to grain refinement techniques it
has been realized that a steel material of smaller grain
size is essential in order to allow the steel material tofeature improved wear strength. NTN has developed
the “FA process”, which is a heat treatment processcapable of reducing the grain size of a bearing steel
material to a level less than half of the previously
possible level (Photo 1). NTN’s FA process is uniquein that the bearing steel, which has already undergone**Automotive Engineering Dept. Automotive Sales Headquarters**Needle Roller Bearing Engineering Dept. Automotive Sales Headquarters-86-Transmission Technology Trends and Product Developments

a conventional carbonitriding process, is further
subjected to a special heat treatment process so thatthe bearing steel material can more positively
withstand surface damage.
Information about the result of test for NTN’s new
products is provided below.

2.1 Rolling contact fatigue life test for ball bearing
under contaminated lubrication condition
The test rig used is schematically illustrated in Fig.
1, and the test conditions applied are summarized inTable 1. The grain size of the contaminant powder(hard material) falls in a range of 100-180 μm, which
is a relatively large size.
The result of life test is summarized in Table 2. The
L10 life of the bearing treated with the FA process is
3.7 times longer than that of the bearing treated withordinary quenching and 2.1 times longer than that of
the bearing having undergone carbonitriding. It waslearned that grain refinement is effective against a
failure mode that results from stress concentrationsuch as flaking caused by dents.
So far, it has been reported that the life of a bearing
steel material in a contaminated lubrication situation
can last longer when the amount of residual austenite
in the material is larger and the hardness of thematerial is greater. As summarized in Table 3, theamount of residual austenite is smaller with the FAtreated material when compared with the carbonitrided
material; however, the life of the FA-treated material islonger. It appears that grain refinement is effective incompensating for the loss in bearing life resulting froma decrease in residual austenite.

0.05mmOrdinary quenched material

Load-applying
coil spring
Load-applying
ball bearing 6312
Test bearing
6206

0.05mmCarbonitrided material

Drive pulley

Test bearing
6206

CouplerFig. 1 NTN rolling contact fatigue test rig for ball bearing
Table 1 Test condition of ball bearing 6206 undercontaminated lubrication

0.05mmFA-treated material
Photo 1 The prior austenite grain boundaries-87-

Load Fr (kN)

6.86

Contact pressure Pmax (GPa)

3.2

Speed (min-1)

3000 (inner ring)

Lubrication system

Turbine 56
Oil bath lubrication, approx. 30 mL of oil

Concentration of contaminant

0.4g/LType of contaminant

Gas atomized powder
Grain size: 100-180 mm
Hardness: approx. HV800

NTN TECHNICAL REVIEW No.75（2007）

Table 2 RCF (rolling contact fatigue) -life test results of 6206 under contaminated lubrication
Heat-treatment
technique

L10 life(h)

L50 life(h)

L10 life ratio(taking the life of material having undergone
ordinary quenching as 1.0)

L10 life ratio(taking the life of material having
undergone carbonitriding as 1.0)

Ordinary quenching

13.1

19.4

1.0

0.6

Carbonitriding

23.0

45.5

1.8

1.0

FA process

48.0

87.2

3.7

2.1

Table 3 Metallurgical properties of ball bearing 6206 (0.05mm depth from surface)Heat-treatment
technique

Mean grain size on conventional
austenite boundary (μm)

HVhardness

Residual stress(MPa)

Residual austenite(%)

Ordinary quenching

10.5

746

-151

7.1

Carbonitriding

9.4

752

-211

25.5

FA process

4.4

733

-233

18.9

L10 life of the bearing treated with the FA process is
4.1 times longer than that of the bearing treated withordinary quenching and 2.0 times longer than that of
the bearing having undergone carbonitriding, in otherwords, the effect of grain refinement is apparent.
Grain refinement can provide longer bearing life bymaking the bearing steel more robust against flaking
caused by dents, regardless of size of thecontaminant particles and bearing shape.

2. 2 Rolling contact fatigue life test for tapered roller
bearing 30206 under contaminated lubrication condition
The test rig used is schematically illustrated in Fig.
2, and the test conditions applied are summarized inTable 4. The contaminant powder (hard material)
comprises 90wt% of smaller grains whose size
measures 50 mm or smaller and 10wt% of larger
grains whose size falls in a range of 100-180μm.
The result of life test is summarized in Table 5. The

SpringTest bearing
PulleyTable 4 Test condition of tapered roller bearing 30206
under contaminated lubrication

Support bearing
Test bearing

Load (kN)

Fr

17.64

Fa

1.5

Contact pressure Pmax (GPa)

2.5

Running speed (min-1)

2000 (inner ring)

Lubrication system

Turbine 56
Oil bath lubrication, approx. 30 mL of oil

Concentration of contaminant

1.0 g/LType of contaminant

Gas atomized powder
50 mm or smaller: 90wt%
100-180 mm: 10wt%
Hardness: approx. HV800

Fig. 2 NTN rolling contact fatigue test rig for tapered
roller bearing

Table 5 The RCF (rolling contact fatigue)-life test results of 30206 tapered roller bearing under contaminated lubrication
Heat-treatment
technique

L10 life(h)

L50 life(h)

L10 life ratio(taking the life of material having undergone
ordinary quenching as 1.0)

L10 life ratio(taking the life of material having
undergone carbonitriding as 1.0)

Ordinary quenching

101.2

117.3

1.0

0.5

Carbonitriding

211.6

284.5

2.1

1.0

FA process

415.6

464.3

4.1

2.0

Table 6 Metallurgical properties of tapered roller bearing 30206 (0.05mm depth from surface)Heat-treatment
technique

Mean grain size on conventional
austenite boundary (μm)

HVhardness

Residual stress(MPa)

Ordinary quenching

11.8

792

-107

6.0

Carbonitriding

12.2

763

-265

32.4

FA process

5.2

748

-258

23.3

-88-

Residual austenite(%)

Transmission Technology Trends and Product Developments

circle diameter leads to a significantly lower torquethough with a slight loss in rigidity. Usually, a smaller
pitch circle diameter means a decreased number ofrollers. However, by adopting a high load capacity
design technique, it is possible to maintain the rigidity
equivalent to that of conventional bearings without adecrease or increase the number of rollers.
Furthermore, by adopting a low-torque cage design, it
is possible to decrease the stirring drag of oil andshear resistance of the oil between the cage bars androllers greatly reducing the torque on the bearing.

3. Up-graded tapered roller bearing
In order to achieve improved fuel economy thedemands are increasing for compact, lightweight, lowtorque designs of tapered roller bearings used intransmissions and differential gearings. To address
this challenge, NTN has already achieved thefollowing improvements for its tapered roller bearing
products:[1] FA process: Compact, light-weight bearing design
has been achieved by seeking longer bearing life
through the adoption of metal material
strengthening based on grain refinement.
[2] High load capacity design: A larger pitch circle
diameter for the cage allowed a maximum numberof rollers to be incorporated, so as to increase theload rating and rigidity of the bearing. This feature
has helped achieve a compact, lightweight bearing
design.
[3] Low torque cage design: By reducing the stirringinduced drag on the oil, the running torque of the
bearing is decreased resulting in improved oil flow.
Also, through a special shape of the cage pockets
the shear resistance of the oil between the cageand rollers is reduced.
By combining these techniques, NTN has
succeeded in developing the “high-rigidity, ultra-low
torque tapered roller bearing” that boasts a lower
running torque without a penalty of shorter life andlower rigidity. Since loss in rigidity on a bearing used
in a transmission or differential adversely affects
meshing of gears, a low-torque bearing not associatedwith loss in rigidity certainly satisfies the market
needs.
Incidentally, to be able to improve fuel economy,
lower viscosity lubricating oil is increasingly used intransmissions and differentials. However, use of lowviscosity oil can lead to poor lubrication can lead tosurface damage failures such as seizure. To address
this problem, NTN has already developed the “micro
HL tapered roller bearing”.
This product is described in detail in the associated“Product Introduction” page.

3. 2 Design study for high-rigidity, ultra-low
torque tapered roller bearing
An example of low-torque design study for 30306D,
which is currently used for a differential pinion
application, is illustrated in Fig. 3. The technical data
for the conventional and low-torque bearing designs
are summarized in Table 7.
AAs summarized in Table 7, in order to reduce therunning torque while still maintaining the rigidity
equivalent to that of the conventional design, the lowtorque design has two additional rollers despite thesmaller roller pitch circle diameter. By compensating
for the loss in dynamic load rating with the lifeextending effect derived from the FA process, the lowtorque design boasts the life equivalent equal to or
better than that of the conventional design.

Conventional design

Low-torque design
16

20.75

10.4

16

φ60

Downsizing of
bearing ispossible.

φ30

φ72

19

φ30

14

Fig. 3 The example of low torque bearing examination

3. 1 Technique for achieving lower torque,
while maintaining rigidity

Table 7 Bearing internal design

A design technique that helps achieve a compact,
low-torque bearing without developing loss in rigidity ispossible only when a high load capacity design isadopted.
To be able to attain a low-torque bearing design,
first we had to analyze the contribution of each design
parameter for tapered roller bearing to a low-torquedesign. Therefore, we learned that a smaller pitch-89-

Conventional
design

Low-torquedesignDynamic load rating (kN)

Cr =49.0

Cr =33.0

Static load rating (kN)

C0r =52.5

C0r =35.5

Roller pitch circle diameter (mm)φ51.54

φ44.44

Number of rollers

15

17

Mass (kg)

0.393

0.223

NTN TECHNICAL REVIEW No.75（2007）

3. 3 Performance verification for high-rigidity,
ultra-low torque tapered roller bearing

4. Efforts for improved functions forplanetary gearing

The running torque measurement results of the lowtorque design and conventional design are plotted inFig. 4. The results of the axial elastic displacementwith these designs are shown in Fig. 5. As shown in
Fig. 4, the new tapered roller bearing design attained
a 50% torque reduction at a range of 2000-3000 min-1
which is a normal speed range for differential pinions.
Additionally, as can be noted from Fig. 5, the new
design boasts rigidity virtually equivalent to that of theconventional design.

4. 1 Optimized high-speed bearing for planetary
gearing
As the planetary gearing runs at a higher speed,stronger centrifugal forces will act on the planet gears
and rollers; as a result, while the bearing is running
the rollers can repeatedly hit the cage, possibly
damaging the cage. To address this problem, NTN
has performed FEM analysis for the bearing as well as
dynamic behavior analysis with the bearing
experiencing planetary motion, and then optimized thecage that is used in a high-speed application.

300

Running torque N・cm

250

Low-torque design

4. 1. 1 Stress analysis for cage (static analysis)

Conventional design

The most often failed points on cages used inplanetary gearing are corner R sections on cage
pockets. To prevent the damaged cage, it isnecessary to mitigate the stress that may occur onsuch a corner. Therefore, fully utilizing our FEM
analysis technique to optimize the shape of the cage,
we have successfully reduced the stress that may
occur at these points by 50% (see Table 8).

200

50%
reduction

150
100

50
0

0

1000

2000

3000

4000

5000

Speed min-1

Table 8 Embodiment of FEM analysis

Axial elastic displacement, μm

Fig. 4 Relationship between speed and torque ofcurrent bearing and low torque bearing

Conventional design

New design

60

Corner R

Corner R

50
4030

4. 1. 2 Stress analysis for cage (dynamic analysis)
We have attempted a dynamic behavior analysis forthe cage subjected to planetary motion that has been
optimized through FEM analysis. The analysis result
for the surface pressure acting between the outer race
(planet gear bore) and rollers is plotted in Fig. 6, andthe stress working on the cage is shown in Fig. 7.
From these results it's apparent that the stressworking on the cage peaks after the contact surfacepressure between the outer race and rollers has
peaked. As a result, we learned that the mechanicalstrength of the cage is most significantly affected by
the roller exiting a loading region and hitting the cage.
For detailed information about our dynamic analysis,refer to the Technical Paper “Dynamic Analysis for
Needle Roller Bearings Under Planetary Motion” inthis issue of NTN Technical Report.

20
Low-torque design
10

Conventional design

0
0

5000

10000

15000

20000

25000

30000

Axial load, N

Fig. 5 Deformation of axial direction-90-

Transmission Technology Trends and Product Developments

4. 2 Improved durability for gear tooth face
One typical example of a gear failure mode ispitting. When pitting arises abnormal noise andvibration will occur on the gearing and the efficiency inpower transmission on the gearing will decrease.
When a gear is running at a higher speed, oil film
forming capability on the gear tooth can bejeopardized due to an increased temperature andmetal-to-metal contact between the tooth face,possibly resulting in pitting.
NTN has adopted a special surface treatment
technique to planetary gears, which can form an oilfilm to achieve longer life and be resistant to peeling(surface flaking failure). Pitting on a gear tooth face isa surface initiated flaking failure (comparable topeeling on bearings). To prevent gear tooth pitting andextend life, NTN has applied its special surface
treatment to the gear tooth faces of its planetary gear
products. An image of the surface having undergone
the special surface treatment is shown in Fig. 8. As
shown here, the specially treated surface has dimples(oil pots) that help improve oil film forming capability.

Surface pressure between outer race and rollers

StressContact surface pressure

Stress on cage

→Time

Maximum stressFig. 6 Relation between outer race-roller surfacepressure and cage stress

Conventional design
New design
Fatigue limit of cage

24000

26000

28000

30000

32000

→Rotation speed of planet gears min-1

Fig. 7 The maximum value of cage stressFig. 8 Image of special surface treatment

4. 1. 3 Function evaluation

4. 2. 1 Improved pitting resistance for gear tooth face

To check relevance of our dynamic analysis, twotypes of cages were incorporated into a planetarytransmission, and then the cages were tested under
the conditions that are the same as those of the
dynamic analysis. The result of evaluation issummarized in Table 9. As the stress applied during
the analysis exceeded the fatigue limit of theconventional cage, all the test pieces of theconventional cage failed. In contrast, none of the test
pieces of the new cage design failed, demonstrating
reliability of our new cage design.

The combinations of the test gears used are
specified in Table 10. Comparison was made in termsof the number of load cycles applied before pitting
failure mode would occur. The example whose gear
underwent the special surface treatment boasts pitting
resistance four times greater than that of the otherexample. Thus, it has been verified that formation ofoil film is effective to prevent pitting on gear tooth face(Fig. 9) as no pitting occurred on the gear that
underwent the special surface treatment.
Incidentally, to achieve planetary gear improvement
NTN has incorporated the above-mentioned elements
into its planetary gear assembly (Photo 2).

Table 9 Evaluation results of cage intensity
Condition

1

2

3

Rotation speed of pinion(min-1)

25000

28000

30000

Conventional
State of cage
designhaving undergoneNew
the test
designFailedNo problem(further operation is possible)-91-

NTN TECHNICAL REVIEW No.75（2007）

number of gear speeds on AT is described in detail in
the relevant “Product Introduction” page.

Table 10 Combination of the gears for examination
Gear!

Gear@

1

Not treated

Not treated2

Not treatedSpecial surface treatment

5. 1 Features of the bearing--lower cross-section(compact design)
The bearing race consists of a thin sheet material
whose thickness measures 0.5 mm. Its staked
sections are shaped such that the thickness reduction
is minimized while sufficient mechanical strength ismaintained on these sections. The cage comprises ofa single sheet metal piece formed by a pressing
process that is capable of accepting a 1 mm diameterroller (Fig. 10).
This cage is held by the staking areas of the innerand outer races to achieve the triple unit design. Notethat the NTN triple unit thrust needle roller bearing
design with 2 mm cross-sectional height is thethinnest design of its kind in the world.

Ratio of loading cycles
before occurrence of pitting

6
5
4
3

Number of rotations:3500mm-1
Toque :120N・m
Lubricating oil :ATF
Oil temperature :80˚C

2
1
0

1

2

Fig. 9 Evaluation results of gear pitching

Bearing race :0.5mm
Roller diameter :1.0mm
Bearing ring :0.5mm
Staking member

Cross-section
height:2.0mmFig. 10 Construction of triple-unit 2mm cross section
thrust roller bearing

5. 2 Cage structure
We have invented a cage of a unique structure to
hold 1 mm diameter needle rollers (Fig. 11). The
single piece cage ensures a sufficient sheet thickness,while the rollers are retained with an ironing
technique. The grooves are formed on the entire cage
circumference in order to promote oil flow and simplify
the ironing process.

Photo 2 Planetary gear assembly

5. Triple unit 2 mm cross-section
thrust needle roller bearing
To help improve fuel economy of automobiles byreducing the torque on AT, sliding bearings have beensuperseded with rolling bearings. At the same time,
there has been a mounting need for a smaller crosssectional height for thrust bearings in order tominimize increasing axial dimension of an AT that
results from the demand of an increased number ofgear sets. Furthermore, a need is also mounting for a
triple unit type bearing having a bearing race that canbe handled easily and is built into an automatictransmission through a smaller number of work steps.
To satisfy these needs, we have developed the“triple unit 2 mm cross-section thrust needle roller
bearing” that, though having a bearing race, boasts a
minimized cross-sectional height.
The NTN high-speed thrust needle roller bearing
capable of higher speeds to cope with an increased

Groove formed on
entire circumference

Roller retaining section

Groove formed on
entire circumference

Roller retaining section

Single-plate cage

Fig. 11 Cross section of new shape pressed-cage

5. 3 Function evaluation
The needle rollers being built into the NTN low
cross-section thrust needle roller bearing are optimallycrowned. The functions (torque measurement, life
evaluation) of the optimized needle rollers have been
compared with those of the conventional needle
rollers having standard crowning.
-92-

Transmission Technology Trends and Product Developments

5. 3. 1 Torque measurement

6. Conclusion

The results of torque measurements with theconventional and new designs are illustrated in Fig.
12. The rotating torque of the new design, low crosssection, optimally crowned, thrust needle roller
bearing is 37% lower compared with the conventional
design.

We are expecting that the working environments forbearings in an automotive transmission will become
increasingly severe. NTN remains committed to thedevelopment and marketing of new automotive
products that can cope with these working
environments and market demands, positively
contributing to mitigation of environmental impacts
imposed by automobiles.

5. 3. 2 Life evaluation
Life test results with the conventional and new
bearing designs are plotted in Fig. 13.
The samples from conventional and new bearing
designs were evaluated under misalignment
conditions using assumptions for automotivetransmissions. As a result, it has been found that our
new design, a low cross-section thrust needle roller
bearing whose rollers have been optimally crowned,
boasts a bearing life 6.3 times longer when comparedwith the conventional design with needle rollers ofstandard crowning.

References
1) Chikara Ooki, Kikuo Maeda, Hirokazu Nakashima:
NTN Technical Review, No.71 pp.2-7 (2003)
2) Takashi Tsujimoto, Jiro Mochizuki: NTN Technical
Review, No.73 pp.30-39 (2005)

Test conditions
Speed :6000min-1
Load :3200N (0.37Ca)
Magnitude of misalignment :5/1000deg
Lubrication system :ATF, 100˚C, circulating lubrication
at 50 mL/min

Test conditions
Speed :1000min-1
Load :0.1Ca
Lubrication system :ATF, 100˚C, circulating lubrication
at 50 mL/min

Running torque N・cm

Running torque N・cm

10
8
6
4

2
0

Conventional design

95
90
80
70
60
50
40
30
20

New design
Conventional design

10
5
7 1
100

New design

Fig. 12 Results of rotating torque

2

3

5 7 1
101

2 3

5 7 1
102

Fig. 13 Results of life test

Photos of authors

Takahiro KANAMOTO

Takashi UENO

Akihiko KATAYAMA

Masanori SATOU

Automotive Engineering Dept.
Automotive Sales Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters

Needle Roller Bearing
Engineering Dept.
Automotive Sales Headquarters

Needle Roller Bearing
Engineering Dept.
Automotive Sales Headquarters

-93-

2

3

5 7 1
103

NTN TECHNICAL REVIEW No.75（2007）[ Technical Paper ]Dynamic Analysis for Needle Roller Bearings
Under Planetary Motion

Tomoya SAKAGUCHI*
A dynamic analysis tool for needle roller bearings in planetary gearsystems has been developed. This tool uses two-dimensional analysis,
considering three degrees of freedom for of both rollers and cage and two
translational degrees of freedom for the planet gears on the radial plane. In
addition, elastic deformation of the cage can be simulated using a
Component-Mode-Synthesis method in order to evaluate cage stress.
Numerical results on the maximum principal stress of the cage using thisdynamic analysis tool indicates that the effect of the planet gear orbital
rotation speed on the cage stress is higher than that of the planet gearrotation speed and the maximum cage stress is in nearly proportional to thesquare of the orbital rotation speed. The reason is that stress on the base of
the cage bar increases in the unloaded zone due to supporting the roller
centrifugal force induced by the orbital motion of the planet gear. In addition, relatively high stresses of the cagewere observed in the two following cases: the roller just passing the load zone and accelerates due to centrifugal
force and collides with the front bar, and cage bar contact force due to a roller in the load zone opposes thecage moments generated by rollers in the non-load zone.

author developed, on general-purpose mechanism
analysis software, a dynamic analysis tool for atapered roller bearing that is capable of considerations
about six degrees of freedom of motion of the cageand rollers as well as elastic deformation of the cage2).
The author previously reported that the trajectory andstress output from this tool coincide well with the test
result 2).
The above-mentioned dynamic analysis tool fortapered roller bearings has been applied to needle
roller bearings under planetary motion. The stressacting on bearing cages was analyzed; wherein, thedynamic analysis was limited to three degrees of
freedom on the radial plane of bearing in order toattain better analysis efficiency.

1. Introduction
A rolling bearing consists of a pair of bearing rings,
a number of rolling elements situated between the
bearing rings, and a cage that separates the rolling
elements at equal intervals. A cage will not fail in an
ordinary application; however, a cage can fail in an
application that accompanies a varying load orplanetary motion. To prevent a failure, a cage must be
provided with greater mechanical strength;
unfortunately, determination of forces acting on a cage
through a test is difficult. Therefore, a dynamic
analysis technique for a rolling bearing which
considers cage behavior will offer a useful solution tothis challenge.
ADORE 1) by Gupta is a known dynamic analysis
technique for rolling bearings that positively reflects
cage motions but, the cage assumed in this techniqueneeds to be a simple-shaped solid body. Furthermore,this technique is not capable of directly determining a
cage stress. If we attempt to determine a cage stressthrough FEM-based structural analysis, it will bedifficult to rationally define the load and supporting
conditions for the cage. To address this issue, the2. Dynamic analysis model
The planetary gear mechanism to be analyzed isschematically illustrated in Fig. 1. There are sevenassumptions about his model, wherein, the model canbe simplified as shown in Fig. 2 due to theseassumptions.

*Elemental Technological R&D Center-94-

Dynamic Analysis for Needle Roller Bearings Under Planetary Motion

transmission torque (Fig. 2).
6. The cage is assumed to be an elastic body basedon the Component-Mode-Synthesis method 4) inorder to take into account its elastic deformation. In
other words, the elastic deformation of the cage isexpressed as the sum of the elastic deformation in
inherent deformation mode and that in restrained
deformation mode. All other members are each
assumed to be a solid member; and local elastic
contact is considered for regions where interference
can occur owing to their geometrical shapes.
7. Generally, three or more planet gears are situatedon the carrier. Assuming that all the planet gears
are equivalent with each other, one planet gear onlyis analyzed. For convenience of analysis, theweight of planet gears is ignored.
Fig. 1 Actual forces and motions on a planet gear

Based on these assumptions, it will be possible, likewith a conventional dynamic analysis procedure forbearings 2), to analyze needle bearings in a dynamic
planet gear. This analysis is done by describing thecontact forces, frictional forces and rolling viscosity
resistances occurring between the rollers and raceway
and the rollers and cage on the general-purpose
mechanism analysis software “MSC.Adams” 5).
To be able to assess these interference forces, theauthor has again used a model for the 2D dynamic
analysis for cylindrical roller bearings 3). Despitechange in the number of dimensions, the basic
concept for tapered roller bearings 2) is same as thatfor the interference model in the present study.
The cage as an elastic body was treated in amanner essentially same as that for a cage in atapered roller bearing 2). However, for the current
analysis, a 2D cage model was developed: as
illustrated in Fig. 3. A 3D geometrical cage model was
cut at the middle cross-sectional plane of the bearing
and constaints for the 2D plane were set up on thisplane. Furthermore, the boundary point in theconstraint deformation mode per the ComponentMode-Synthesis method 4) was situated at the
midpoint of each bar of the cage on the above
mentioned cross sectional plane. The cage mass on
the general-purpose mechanism analysis software
package is half as that of the actual cage. To copewith this situation, the other constituting elements
were cut at their center in the axial direction so thatthe mass and inertial moment of each constituting
element are half those on the actual cage.
Accordingly, the magnitude of the interference force,such as a contact force or frictional force, acting
between these elements was assumed to be a half theactual forces.
It was assumed that the interference force from a

Fig. 2 Analyzed forces and motions on a planet gear

1. The analysis technique adopted is a twodimensional dynamic analysis where degrees of
freedom are limited only within a radial plane.
2. Apparent forces including the centrifugal force aretaken into account.
3. The carrier rotates at a constant speed about afixed axis.
4. The outer ring raceway (planet gears) has twotranslational degrees of freedom, provided that therotation speed of the planet gears is known.
5. It is assumed that the radial internal clearance on
the bearings is smaller than the mesh clearance on
the gearing and all the centrifugal force on theplanet gears acts on the rolling bearings. Morespecifically, it is assumed that the sum ofinterference forces from the sun gear and that fromthe ring gear each acting on the planet gears is
equivalent only to the reaction force of the-95-

NTN TECHNICAL REVIEW No.75（2007）

roller acts on a corresponding pocket on the contact
surface at the center of the rib. The maximum major
cage stress often occurs at the base of a bar of cage
pocket, and as a result, the above-mentioned
assumption tends to slightly overestimate the stress.
Since the cage in question is guided by the outerring raceway, the contact force and frictional force
acting between the cage and the outer ring were
evaluated. This contact force was calculated at threeparticular points along the cage outer surface, foreach cage pocket, as shown in Fig. 4. Each contact
force value was calculated based on the amount ofgeometrical indentation into the outer ring raceway.
The elastic deformation characteristics of the cagewere calculated using, the “I-deas NX Series*1” (UGS)6),
which is an FEM software package. For dynamic
calculation, a mechanism analysis software package,“MSC.Adams” 5), as well as an optional package,“ADAMS/Flex” 5), which is an optional variant of
MSC.Adams, were used. For stress evaluation,“ADAMS/Durability” 5), which is another optionalvariant of MSC.Adams, was used.

3. Examples of analysis
The technical data for the planetary gear system
and needle roller bearings used for the present study
are summarized in Table 1. The cage used was anouter ring guided steel cage (model KMJ-S) fabricatedthrough a press-forming and welding process.

Operatingconditions

Bearing
Planetarytechnical data gear system

Table 1 Specifications of needle roller bearing andplanetary gear system
Pitch circle diameter of sun gear, mm
Pitch circle diameter of planet gears, mm
Pitch circle diameter of ring gear, mm
Mass of planet gears, kg

184.2
57.4
299.0
0.045

Outer ring raceway dia., mm
Inner ring raceway dia., mm
Roller dia., mm
Roller length, mm
Number of rollers

19.85
13.85
2.997
13.8
11

Orbital rotation speed of planet gears Nc, min-1
Rotation speed of planet gears Np, min-1
Torque transmitted per planet gear T, N･m
Lubricating oil
Typical temperature of lubricating oil, ˚C

5940
25000
16, 32
ISO VG100
120

4. Stress generating mechanism of the cage
The analysis result shown in Fig. 5 was obtained at
the instant where the orbital rotation speed of theplanet gears was at 5940 min-1 and the major cagestress was at the maximum, wherein the direction of
motion of the related components is same as thatshown in Fig. 2. In Fig. 5, the centrifugal force is in
the upward direction. The maximum major stressoccurred at the base of pocket bar when a roller in the
non-load zone came into contact with this bar. The
contact force triggering this stress was induced by thecentrifugal force on the roller.
To be able to review the state of contact between
the rollers and cage on a running bearing, themaximum stress history is illustrated in Fig. 6. The
stress values presented are non-dimensional values
obtained by dividing by the fatigue strength of thecage material. To be able to locate the load zone, thedisplacements on the adjacent rollers along the radial
direction of the planetary gear system are shown inthis graphical representation. The above-mentioned
maximum stress occurs because a cage bar exiting
the load zone will be subjected to the centrifugal force
of a roller in the non-load zone, thereby the a tensilestress and a compression stress sequentially occur inthe right and left non-load zones as shown by 1 and
2 in Fig. 7. This maximum stress occurs at the .0173
point in Fig. 6. In the load zone, the frictional force

Fig. 3 Cage elastic model introduced into twodimensional dynamic analysis

Fig. 4 Locations of interaction forces on cage outside*1: The I-deas NX Series is a trademark or registeredtrademark of UGS Corp. or its subsidiary in USA andother nations.

-96-

Dynamic Analysis for Needle Roller Bearings Under Planetary Motion

Dimensionless stress

2.0

2.0

1.0

1.0

0.0

0.0

-1.0

-1.0

-2.0
0.015

0.01750.02

0.0225-2.0
0.025

Dimensionless displacementon the cage is generated due to the frictional force
between the cage outer surface and the outer ring
raceway as well as the centrifugal force on the roller in
the non-load zone.
In Fig. 6, a relatively large peak in tensile stress ispresent immediately after exiting the load zone. This isbecause around the end of the load zone, the roller issituated at the trailing end of the pocket and after
exiting the load zone, the roller accelerates due to thecentrifugal force acting on it and hits the bar in front ofit (4 in Fig. 7).

occurring from the raceway is greater than thecentrifugal force occurring on the roller, thereby thecentrifugal force on the roller does not act on thecorresponding cage bar. At the 0.0212 s point in Fig.
6, a compressive force appears to be present. Thismeans that in the latter half of the load zone, the bar isin contact with the roller in front of it (3 in Fig. 7), andas a result of this situation, a decelerating moment isexerted onto the cage. This decelerating moment
occurs because an accelerating moment occurs in the
non-load zone on the cage. The accelerating moment

Time sec.

Fig. 6 Stress histories of a cage pocket bar and theadjacent rollers displacements along the radial directionof planetary gear systema) all over the bearingb) local view near the maximum principle stress point
Fig. 5 Analysis result of a needle roller bearing underplanetary motion (Nc: 5940 min-1, T = 32 N･m)Fig. 7 Principal mechanisms to make cage stress rise underplanetary motion-97-

NTN TECHNICAL REVIEW No.75（2007）

torque T was 32 Nm and the orbital rotation speed
was 7000 min-1. This situation was true with thephenomenon 4 in Fig. 7. However, when the orbital
rotation speed was 8000 min-1, a different result was
obtained as the collision shown in the phenomenon 4
in Fig. 7 was less significant partly because the loadzone position shifted due to an increase in thecentrifugal force acting on the planet gears.
Next, the outcome from the reversed transmission
torque direction was studied. Consequently, it has
been found that, at the rotation speed range of 5000
min-1 or lower, the stress at T = –32 Nm is greaterthan that at T = –16 Nm. Additionally, at the rotation
speed range of 6000 min-1 or higher, the stress at T =–32 Nm is smaller than that at T = –16 Nm. In a higher
rotation speed range, and when T= –16 Nm, a stressof greater magnitude occurred from collision in thesituation 4 in Fig. 7. However, when T = –32 Nm, theroller at the trailing end of the load zone has alreadybegun to come into contact with the cage bar in frontof it in the orbital rotation direction of the planet gears.
As a result, the stress was relatively smaller (show in
Fig. 8) due to a smaller difference in relative
speed between the roller and cage car.

5. Effects of the operating conditionsof the planetary gear system onto
the cage stressFig. 8 plots the interrelation between the maximum
cage stress and the orbital rotation speed of the planetgears. Assuming that the magnitude of thetransmission torque T was 32 Nm or 16 Nm, thecalculation result has been plotted. For transmission
torques in the reverse direction, the calculation result
has also been plotted. Though the stress varies
depending on the magnitude and direction oftransmission torques, the cage stress generally
increases with a greater orbital rotation speed of theplanet gears. As discussed in the previous section,
there is a close interrelation between the centrifugal
force on a roller and the cage stress. In Fig. 8, thestress is roughly in proportion with the square of anorbital rotation speed of planet gears.
A greater stress occurred when the transmission

2.5

T : 32Dimensionless stress

2.0

T : 16
T : -32

1.5

T : -16
1.0

0.5

0.0
0

2000

4000

6000

8000

10000

Planet gear orbital rotation speed Nc, min-1

Fig. 8 Cage stress with various speeds of planet gear orbital rotationand transmitted torque2.5

Np : +, T : 32
Dimensionless stress

2.0

Np : -, T : 32
Np : +, T : -32

1.5

Np : -, T : -32
1.0

0.5

0.0
0

20000

40000

60000

80000

Planet gear rotation speed Np, min-1

Fig. 9 Cage stress with various speeds of planet gear rotation andtransmitted torque-98-

As can be derived from Fig. 8, the factor
most apparently affecting the cage stress at
a given planet gear orbital rotation speed is acollision phenomenon (4 in Fig. 7) between
the roller immediately after exiting the loadzone and the corresponding cage bar.
Assuming that the magnitude of thetransmission torque onto the maximum cagestress could be ignored, the result at
approximately 7000 min-1 was considered.
Consequently, it was learned that the stressat T = 32 where collision occurred is 100%
greater compared with that at T = 16; andthat the stress at T = –16 is 70% greatercompared with that at T = –32. In otherwords, it may be concluded that the
maximum cage stress when collisionphenomenon has occurred is approximately
85% greater compared with that whencollision phenomenon has not occurred.
The orbital rotational speed of the planetgears was fixed at 5940 min-1 and the nondimensionalized maximum cage stress wasmeasured while varying the speed anddirection of rotation of the planet gears andthe direction of transmission torques. The
resultant maximum stress values are plottedin Fig. 9. Under all the conditions, the cagestress relative to the rotation speed of theDynamic Analysis for Needle Roller Bearings Under Planetary Motion

acting between the constituting elements of needle
roller bearings as well as the resultant cage stress.
The orbital rotation speed of the planet gears hadgreater affect on the cage maximum stress comparedwith the rotation speed of the planet gears and isnearly proportional to the square of the orbital rotation
speed. This is because the orbital rotational motion of
the planet gears causes the centrifugal force to occur
on the rollers. This centrifugal force acts on thecorresponding cage bar, in the non-load zone, and thestress increases at the base of this cage bar.
Furthermore, a relatively large cage stress wasobserved if the roller at the trailing end of the loadzone is situated in the rear of the cage pocket
opposite the direction of rotation when the roller
exiting the load zone is accelerated by the centrifugal
force hitting the cage bar in front of it. Large stresses
also occur if the moment acting on the cage in the
non-load zone is balanced by the moment resulting
from the contact force between the roller and cage barin the load zone.

planet gears is proportional to the speed raised to the
0.34th power. The increase in the cage stressresulting from the increase in the rotation speed stems
from the increase in the difference of the relative
speed between the rollers and the cage.
From Figs. 8 and 9, it would be understood that theorbital rotation speed of the planet gears significantly
affects the cage stress. This trend, as discussed in theprevious section, stems from the centrifugal force on
the rollers. At the same time, the cage stress will vary
when the direction and magnitude of the transmission
torque as well as the speed and direction of rotation of
the planet gears are varied. The cage stress canrapidly increase particularly when the roller at thetrailing end of the load zone is situated at the back of
the cage pocket and this roller exits the load zone andis accelerated by the centrifugal force thereby hitting
the cage bar in front of it. If this situation takes place,the cage stress can increase by approximately 85%
as compared with a case where no collision has
occurred.
The operating conditions under which the collisionphenomenon discussed above appears to be
determined by conditions including the angle of roller
orbital motion at the load zone exit as well as themoment acting on the cage in the non-load zone.
However, the affects of these operating conditions
need to be detailed in the future.

References
1) Gupta, P. K.: Advanced Dynamics of Rolling
Elements, Springer-Verlag, New York (1984).
2) Sakaguchi, T., and Harada, K.: "Dynamic Analysis ofCage Stress in Tapered Roller Bearings," Proc.
ASIATRIB 2006 Kanazawa, Japan, (2006)649-650.
3) Tomoya Sakaguchi, Kaoru Ueno: Dynamic Analysis ofCage Behavior in a Cylindrical Roller Bearing, NTN
Technical Review, No.71 pp.8-17 (2003)
4) Craig, R. R., and Bampton, M. C. C.: "Coupling
substructures for dynamic analysis," AIAA J., 6 (7),pp.1313-1319 (1986)
5) MSC.Adams (Registered trademark of MSC.Software
Corporation)
http://www.mscsoftware.co.jp/products/adams/6) I-DEAS: http://www.ugs.jp/product/nx/I-deas.html

6. Conclusion
A 2D dynamic analysis tool for needle roller
bearings under a planetary motion has beendeveloped. This tool allows its users to take into
account the elastic deformation of a cage whosemechanical strength tends to be low relative to theconstituting elements within a rolling bearing. The
users can calculate the resultant cage stress based on
the magnitude of this elastic deformation. Thus, thistool has been used to study the interference force

Photo of author

Tomoya SAKAGUCHI
Elemental Technological
R&D Center-99-

NTN TECHNICAL REVIEW No.75（2007）[ New Product ]Micro HL Tapered Roller Bearing

Takashi UENO*
With the viscosity of the oils being used in automotive transmissions
and differentials becoming lower, oil film parameter lambda is falling
below 0.5 (VG10 oil used). In order to handle such increasingly severe
lubrication conditions, NTN has developed Micro HL (high-lubrication)
tapered roller bearings.
Micro HL treatment forms a finer dimple in the metal thanconventional HL treatments and maintains surface roughness equal toa super-finish. Thanks to the micro oil pot effect, adequate oil films canform even under severe lubrication conditions of lambda 0.2 to 0.5,
conditions in which it’s difficult for oil film to form. Therefore, thisprocess maintains calculated service life by reducing lubrication
limitations.

To address the need for a much lower lubricant
viscosity in the future, NTN has developed the “micro
HL (High-Lubrication) tapered roller bearing” that isessentially a tapered roller bearing incorporating
optimized HL treated rollers. Fig. 1 summarizes therelation between the oil film parameter (described laterin the report) and the life ratio relative to the standard
bearing. As shown in Fig. 1, when the oil film
parameter is 0.2 or lower the life of the HL bearing
samples are shorter than those of the standard
bearing samples. In contrast, even when the oil film
parameter is 0.5 or lower (a very severe lubricating
condition), the micro HL tapered roller bearing
samples boast better oil film forming capability and

1. Introduction
Recently, lower viscosity oil use has increased in
automotive transmissions and differentials in order todecrease fuel consumption (reduction in CO2emissions) and help reduce power needed for steering
under low temperature conditions. Incidentally, theadoption of smaller and more powerful engines
causes the oil temperature to be higher. This in turnleads to a decreased lubricant viscosity 1).
In such circumstance, more bearings used in
automotive transmissions and differentials will haveincreased failure due to surface damages such as
peeling, smearing and seizure; each failure resulting
from poor oil film formation, rather than internal flaking
that may occur even when oil film is positively formed.
A bearing that fails due to surface damage is shortlived, and does not satisfy the rolling fatigue life of
90% reliability. To address this problem, NTN has
developed the “HL bearing” that is optimized for
operation under low-viscosity lubrication conditions 2).
The “HL bearing” is unique in that the surface of thetapered rollers has fine dimples to retain the oil; thisarrangement helps improve oil film forming capability
for the bearing. Regrettably, when the HL technique isapplied to tapered roller bearings where sliding
contact occurs on the rib portion, the seizure-immunityof the rib portion could be jeopardized.

Life ratio relative to standard bearing

Future oil
Conventional oil
14
12

HL

10
8

Micro HL

6

Standard bearing

4
2
0
0

0.2

0.4

0.6

0.8

1.0

1.2

1.4

Oil film parameter Λ

Fig. 1 Relationship between oil film parameter and life
ratio compared to standard bearing*Automotive Engineering Dept. Automotive Sales Headquarters-100-Micro HL Tapered Roller Bearing

longer life when compared with the conventional HL
bearing samples. The structure, advantages, andresults of the evaluation test for the micro HL tapered
roller bearing are hereunder described.

3. Basic principle of micro HL
As explained above, it is very difficult to obtain agreater oil film parameter. Therefore, NTN hasaddressed this problem by implementing the three
approaches with its micro HL that are described below.

2. Facts about oil film parameter
It is widely known that the life of any bearing
correlates with the oil film parameter that isdetermined by the elastic hydrodynamic lubrication
theory (EHL theory), which is based on the ratio of theoil film thickness on the contact area of the bearing
components to the combined roughness of thesecomponents 3). Therefore, a greater oil film parameterfor the contact area appears to be effective inpreventing various modes of surface damage on arolling bearing. The oil film parameter formula is given
below:
Oil film parameter

Λ＝

hmin

3. 1 Approach based on lubricant flow model
By referring to the lubricant flow model in thecontact area on a bearing 4) in Fig. 2, we describe
NTN’s basic way of addressing the challenges for both
HL and micro HL. Pictures (a) through (c) in Fig. 2
show similar two-dimensional roughness, whereineach picture exhibits a unique directional path with thefinished surface (note difference in three-dimensionalsurface roughness). The contact points in each picture
are unique and these unique contact points affect the
lubricant flow indicated with the broken lines. As also
shown in Fig. 2, the degrees of smoothness of the
lubricant flow can be arranged in the order of (a)smoothest flow, (b) and (c) slowest flow. In contrast,
the longer the time during which the lubricant that hasentered the contact area remains there, the greater
the oil film forming ability of the lubricant. The degreesof ability for oil film forming can be arranged in theorder of (c) greatest ability, (b) and (a) lowest ability.
In other words, even with the same quality of the twodimensional surface roughness, the directional pathon the finished surface (that is, the quality of threedimensional surface roughness) significantly affects
the trend in oil film formation.
Now, let us think of the effect of the directional pathof the finished surface onto the oil film forming
capability of the this surface. NTN’s dual cylinder tester
and test specimen used for this purpose are illustrated
in Fig. 3. Two types of specimens were used—, inwhich the direction of tool marks on one specimen typewas in the circumferential direction (Fig. 2(a)) and theother type was in the axial direction (Fig. 2(c)). As
shown in Fig. 4, the specimens having axial direction
tool marks are capable of improved oil film formationand have better oil film parameter values.

………（1）

σ12＋σ22
where:
hmin : minimum oil film thickness [mm]σ1 : square mean roughness of bearing ring
raceway surface [mm]σ2 : square mean roughness of roller rollingsurface [mm]Typical measures for attaining a greater oil film
parameter include: (1) maintenance of higher viscosityof the lubricant under actual operating conditions and(2) decrease in the combined roughness on the
bearing’s race and rollers. However, as previously
described in Sec. 1, low-viscosity lubricants are
increasingly being used in automotive applications
and the lubricant viscosity will be further lowered
owing to an increased lubricant temperature.
Furthermore, the combined roughness on the bearing
race and rollers cannot be further improved because
the present-day bearing races and rollers mentioned
in this report are already super-finished.

Contact areas（a）

Lubricant flow
（b）

（c）

Fig. 2 Directional characteristics of finished surfaces and their effect on lubricant movement in a flow model-101-

NTN TECHNICAL REVIEW No.75（2007）

Finer dimples
Circumferential
tool marks（Driver）
D cylinder

D cylinder
F cylinder
Axial toolmarks
D cylinder
F cylinder

F cylinder (super-finished cylinder)（Follower）

Metal-to-metal contact percentage %Fig. 3 NTN’s dual cylinder tester and test specimen

Fig. 5 Observation of roller surface with finer dimples

100

Variation in the circumferentiallymachined surface

80

60

1μm
φ40／φ40（R60）
SUJ2

Fig. 6 Cross-sectional view of roller surface trackwith finer dimples

Spindle oil

40

Pmax 2.3GPa
2000min-1
20

0.1mm

Variation in the axiallymachined surface

Lubricant flow

Contact areas

Finer dimples

0
0

1

2

3

4

Time hFig. 4 The influence of finished surface directional
characteristics on oil film formation

3. 2 Approach based on form of dimples onroller surface
Fig. 5 shows a view of the surface of a tapered
roller having the finer dimples, a unique feature on
both HL and micro HL bearing products, which are
represented by the black spots in this picture. A resultof the measured surface of the rollers using a surface
roughness tester is plotted in Fig. 6, where each finer
dimple on the roller surface is represented on the plot
as a valley.
The flow of lubricant over the roller surface can bemodeled as shown in Fig. 7, which is in accordance tothe author’s flow model. The hatched round areas inFig. 7 show the contact zones that occur fromprotrusions and deformations on the rough surface,
while the cross-hatched round areas represent thezones containing finer dimples. The broken lines
represent the direction of lubricant flow. The rolling
direction is from the left side to the right side in the
diagram and the lubricant fluid flow bypasses thecontact point on the flat smooth surface. Furthermore,

Fig. 7 A flow model showing fluid on finer dimples contact
surfacethe amount of fluid flowing increases around the finer
dimples and this increased flow through the contact
surface results in the surface having a better oil film
forming capability, as compared with the conventionalsurface whose finish orientation (tool mark orientation)is in the circumferential direction.

3. 3 Optimal roughness for application totapered roller bearing
There is sliding contact between the roller large end
face and large end rib of the inner race on a tapered
roller bearing. To better resist seizure -for a tapered
roller bearing, the sliding contact surfaces on this typeof bearing are super-finished. However, the quality ofsurface roughness on the HL treated tapered roller-102-Micro HL Tapered Roller Bearing

bearing is inferior to that of the super-finished surface.
As a result, metal-to-metal contact will occur on the HL
treated tapered roller bearing due to the lack of oil film
forming capability when subjected to an oil film
parameter that is 0.5 or less. To cope with thisproblem, the quality of the surface roughness finish
must be equivalent to that of super-finished process.
The micro HL tapered roller bearing is the result of
an improvement based on the three above-mentioned
approaches.

5. Performance of micro HL tapered
roller bearing
To assess the performance of the micro HL tapered
roller bearing, a life test was performed under severe
low viscosity lubrication conditions with an oil film
parameter Λ=0.2 and Λ=0.5 (standard oil film
parameter Λ=approx. 1.5).
The actual test conditions associated with the oilfilm parameter Λ=0.2 are summarized in Table 1, andthe technical data for the bearing specimens tested
are provided in Table 2. Also, the results of the
bearing life test are graphically plotted in Fig. 9. Theselubricating conditions are extremely demanding: as aresult the specimens of the standard bearing failed
owing to the damage starting at the raceway surface

4. Structure of micro HL tapered roller
bearing
Fig. 8 provides a comparison between standard
rollers and micro HL treated rollers in terms of thequality in three-dimensional surface track. The surfaceof the rollers that underwent the micro HL treatment
has finer dimples while still retaining the surface
roughness equivalent to that obtainable from a superfinishing technique.

Table 1 Test condition
Loading conditions Fr=19.1kN，Fa=6.4kN (P/C＝0.41)
Running speed

Ni=2500 min-1，

Lubricant

JOMO HIGH SPEED FLUID VG1.5 (clean oil)

Lubrication system Oil bath (oil level is at the shaft center)
Calculated life

L10＝92.2h

Table 2 Comparison of test bearing spec
Bearing ring

Rollers

Standard Raceway surface: super-finished Rolling contact surface: super-finished
Roller large end face: super-finishedbearing Rib: super-finished
HL

Raceway surface: super-finished
HL treated
Rib: super-finished

Micro HL

Raceway surface: super-finished
Micro HL treated
Rib: super-finishedStandard roller
φ45×φ81×16（mm）

Bearing size

Finer dimples

Dynamic load rating

Cr＝42kN

Static load rating

Cor＝52kN

Data for specimens suc→ cessfully completed thetest, not developing failure
Cumulative percent failed %99

Micro HL treated roller
Fig. 8 Comparison of 3D surface track between
micro HL treated roller and standard rollerStandard bearing

80

HL

50

Micro HL

20
10130 h or longer

5

11.3h
9.6h

1
100

101

Life h102

103

Fig. 9 Bearing life test results under severe lubrication
conditions of lambda 0.2-103-

NTN TECHNICAL REVIEW No.75（2007）

only 9.6 h after the test began, which is much lessthan the calculated life of 92 h. In contrast, the micro
HL tapered roller bearing specimens achieved a life of
130 h or longer which is higher than the calculated life.
Thus, it has been verified that the life of the micro HL
tapered roller bearing is at least 13 times longer thanthat of the standard bearing.

6. Conclusion
The NTN micro HL tapered roller bearing boasts an
improved oil film forming capability even under severe
lubrication conditions, thanks to a unique fine surface
modification technique. Thus, the author believes that
the NTN micro HL tapered roller bearing products will
operate reliably even if they are lubricated with lessviscous oil.
The micro HL tapered roller bearings may be usedin demanding situations including high-speed
applications where poor lubrication can occur. They
can run in poorly lubricated conditions, or in a specificarea of an application where temperature increase
should be avoided. Thus, the micro HL tapered roller
bearing series will be a product that is positively
accepted in the market.

The actual test conditions associated with the oilfilm parameter Λ=0.5 are summarized in Table 3, andthe technical data for the bearings tested is provided
in Table 4. Also, the results of the bearing life test aregraphically plotted in Fig. 10. The micro HL tapered
roller bearing specimens achieved a life of 600 h or
longer, which is much higher than the calculated life of
88 h. Thus, this verifies that the life of the micro HL
tapered roller bearing is at least 5 times longer thanthat of the standard bearing.

References

Table 3 Test condition
Test conditions

Fr=20kN，Fa=10kN (P/C＝0.49)

Running speed

Ni=2500 min-1，

Lubricant

ENEOS Super Oil T10 (VG10 clean oil)

Lubrication system

Oil bath (oil level at the roller bottom end)

Calculated life

L10＝88h1) Toshio Kunugi: Drive Lubricating Oil, Journal ofJapanese Society of Tribologists Vol.39 No.10 p.845(1994)
2) NTN Catalogue: HL Bearings, CAT. No.3020/J
3) Ryuki Yamashita: EHL in Rolling Bearings, Journal ofJapanese Society of Tribologists Vol.49 No.4 p.330(2004)
4) Patir N. & Cheng H.S., "An Average Flow Model for
Determining Effects of Three-Dimensional Roughness
on Partial Hydrodynamic Lubrication," Trans. ASME,
J.Lub.Tech., 100(1978)12.

Table 4 Comparison of test bearing spec
Bearing ring

Rollers

Standard Raceway surface: super-finished Rolling contact surface: super-finished
Roller large end face: super-finishedbearing Rib: super-finished
Micro HL

Raceway surface: super-finished
Micro HL treated
Rib: super-finished
φ40×φ76.2×17.5（mm）

Bearing size
Dynamic load rating

Cr＝46.5kN

Static load rating

Cor＝56kN
Data for specimens suc→ cessfully completed thetest, not developing failure

Cumulative percent failed %99Standard bearing
80

Photo of author

Micro HL

50
20
10103.7h5

1

101

600 h or longer

102

103

104

Life hTakashi UENO

Fig. 10 Bearing life test results under severe lubrication
conditions of lambda 0.5

Automotive Engineering Dept.
Automotive Sales Headquarters-104-

NTN TECHNICAL REVIEW No.75（2007）[ New Product ]High Speed Thrust Needle Roller Bearings

Kosuke OBAYASHI*
NTN has developed a high-speed thrust needle roller bearing capableof 1.5-million dmn to meet the high-speed demand of high-powered
electric motor units used in hybrid vehicles and multiple speed
automatic transmissions.
In recent years, the demand for high-speed needle roller bearings has
increased for high-powered electric motor units used in hybrid vehicles
and multiple speed automatic transmissions in the pursuit of more fuelefficient vehicles.
NTN’s design features the following:¡Improved surface finish of the cage pocket contact with the roller end, while still maintaining a conventional
cage configuration that has good oil flow.
¡Special contact configuration of roller end and cage pocket which does not affect roller movement.
¡Point contact of the roller end and cage pocket at the center point of the roller end with low circumferential
velocity is made possible by the configuration of roller end curvature.
These features enable reduction of cage wear and seizure as well as improve the permissible rotational speed
from the conventional 860,000 dmn to 1.5 million dmn, which is the world’s highest. In addition, the temperaturerise at high-speed operation is reduced by 36% in comparison with conventional bearings while also improving
lubrication reliability.

This paper hereunder describes the features andperformance of this unique high-speed thrust needle
roller bearing.

1. Introduction
Needs for vehicles that boast lower fuelconsumption or better fuel economy have been everincreasing. As a result, smaller motors for use on
hybrid vehicles and automatic transmission systems
with increased number of gear speeds are morecommonly being used in the automotive market, thusincreasing the demand for thrust needle roller
bearings capable of higher speeds.
NTN’s newly developed high-speed thrust needle
roller bearing (Photo 1) is based on the W-shaped
NTN standard cage (Photo 2) that features better
lubricant flow; and has been optimally designed. Thus,
the new high-speed thrust needle roller bearing boasts
an increased maximum allowable bearing speed, andhas a dmn = 1.5 million.

2. Structure of high-speed thrustneedle roller bearing
By special treatment of the outside diameter of thepocket faces on the cage (Table 2) with which theroller end face contacts, we have improved thesurface quality (surface roughness) of the contact
surfaces of the NTN high-speed thrust needle roller
bearing (Tables 1 and 2).
As a result of the special treatment, a larger contact
surface occurs between the outer pocket face of thecage and the roller end face. Thus, a very stable cage
pocket-roller end face contact state is maintained
regardless of the motion of the cage or roller.
In addition, NTN employs special needle rollers
each having a unique curved end face. Consequently,
the center point of each needle roller, at which theroller circumferential velocity is low, can reliably keepa contact point relation with the cage pocket outside
diameter resulting in an arrangement that has a lower*1: dmn value = dm (pitch circle dia. of bearing, mm)×n (bearing speed, min-1)
Ex. : When a bearing of dm=100 mm is run at
n=10,000 min-1, the resultant dmn value is 100×
10,000 = 1 million.

*Needle Roller Bearing Engineering Dept. Automotive Sales Headquarters-105-

NTN TECHNICAL REVIEW No.75（2007）

Photo 1 High-speed Thrust Needle Roller Bearings

Photo 2 Current Bearings

Table 1 Diagrammatic illustration of development
bearing and current bearing

Table 2 Enlarged illustration of development bearing andcurrent bearing
Bearing
specification

Cage pocket outside diameter

New design

Cross section view

New design

Bearing
specification

Contact surface(treated with new technique)

Conventional design

Conventional design

The cage pocket outside diameter is in contact
with the curved roller end face.

The cage pocket outside diameter is in contact
with a portion of the roller end face (flat face).

Contact surface

PV value.
Furthermore, the cage riding clearance between therollers and cage and also the riding position and theroller crowning form have all been optimized toprovide a certain degree of freedom to the rollers, sothat the cage least affects the behavior of each roller.
Thus, we desire that each roller be independent of the
motion of any other component. As a result of these
efforts, premature wear and excessive temperaturerise of the bearing have been minimized and theallowable running speed of the bearing has been
increased, thereby attaining a dmn value of 1.5 million.

3. Advantages of the new NTN highspeed thrust needle roller bearing[1] Fastest in the world
dmn value of 1.5 million (76% improvement over theconventional product)[With a bearing that is run at a speed of 21,000 min-1,
where the bearing pitch circle diameter measures 72mm][2] Inhibition of temperature rise
36% decrease in bearing temperature rise (relative
to the conventional design)[Data obtained from the operation at an oil temperature of 80˚C and a dmn value of 86,000][3] Cage strength
1.3 times stronger (relative to the conventional
design)-106-

High Speed Thrust Needle Roller Bearings(1) Analysis conditions

4. Performance evaluation

The analysis condition is summarized in Fig. 1.
For the analysis conditions, the outer circumference
and bore circumference of the cage have been
assumed to be fully constrained and a linear load has
been assumed to be applied to the roller riding
positions.

In evaluating the performance of the new NTN highspeed thrust needle roller bearing the conventional
design was used as a benchmark. The conventional
design and the new design have been subjected to aseries of tests that included stress analysis, anallowable limiting speed test and a temperature rise
test. A portion of the performance evaluation test
result is presented below.

(2) Analysis result
The analysis result is summarized in Table 4.
The cage bar strength of the new design has been
compared with that of the conventional design. As aresult, it has been found that the ratio of the maximum

4. 1 Technical data for test bearings
The test bearings used for performance evaluation
measure 63mm in bore diameter, 80mm in outside
diameter and 2.5mm in width.
The bearing technical data for the new design andconventional design are summarized in Table 3.

Bearing size (mm)Basic dynamic load rating Ca (N)27800

4. 2 Analysis of cage stress

Basic static load rating C0a (N)143000

Table 3 Comparison of bearing specifications
Bearing specification

The cage stresses on the new design andconventional design have been analyzed through FEM
analysis technique, and the resultant findings are
summarized below.

New design

Conventional design

Bore dia. 63×outside dia. 80×width 2.5

Cage type

Cage of new form
made of sheet steel

W-type cage
made of sheet steel

PCD (mm)

72.0

71.5

Number of rollers

44

44

Fully constrained(outer circumferential face)

Fully constrained(bore circumferential face)

Application of linear load

Fig 1 Analytical condition
Table 4 Analytical result
Maximum mainstress ratio

New design

Location of maximum principal stress occurrence/Corner R of the pocket inner diameter side)

1

Conventional design

Bearing
specification

0.77

-107-

NTN TECHNICAL REVIEW No.75（2007）

principal stress on the new design to that on theconventional design is 1:0.77. This in turn means thatthe cage bar strength of the new design isapproximately 1.3 times stronger then that of theconventional design.
On the new design, the pocket position is closer tothe outer circumference, as compared with that on theconventional design; thereby the cage bar width
between the corner R portions at the inner diameterside of the neighboring cage pockets is greater, witheach cage bar having greater strength. Furthermore,
the width of the cage bar is again larger at the corner
R area on the outside diameter side of the cage
pocket, as not only the cage bars but also the cageouter circumference rib can be easily deformed. The
resulting mechanical strength of the newly developed
cage is approximately 1.5 times greater then that of
the conventional design.
From these two facts, the newly developed NTN
high-speed thrust needle roller bearing boasts a
much-improved mechanical strength of the cage bars,
compared with the conventional product.

The NTN horizontal-type thrust bearing test rig usedis schematically illustrated in Fig. 2.

(1) Test conditions
The conditions used for the allowable limiting speed
test are summarized in Table 5.

(2) Test result
The bearing temperature on the new design isgraphically plotted in Fig. 3. Table 6 provides
appearances of the cage pocket outside diameter faceand roller end face for the tested specimens.
The new design was tested at oil flow rates of 100
mL/min, 200 mL/min and 300 mL/min with nospecimen developing a problem of roller drilling wearon the cage. Thus, it has been verified that our newly
developed high-speed thrust needle roller bearing iscapable of readily running at 21,000 min-1 (dmn = 1.51
million), which is the maximum permissible speed of
the test rig, regardless of the magnitude of oil flow
rate. Thus, this speed was taken as the allowable
limiting speed of this new bearing.
At an oil flow rate of 100 mL/min, the cage on theconventional design developed roller drilling wear andthe end face of the roller exhibited wear when theconventional design was running at 13,000 min-1 (dmn
= 93,000). Therefore, the allowable limiting speed of
the conventional design should be limited to 12,000

4. 3 Allowable limiting speed
For the allowable limiting speed test, an NTN
horizontal-type thrust bearing test rig was used.

Table 5 Limiting speed test condition
Test bearing

Load

Test rig

NTN horizontal-type thrust bearing test rig

Running speed

21000 rpm (dmn = 1.51 million)

Axial load

1961 N

Lubrication
systemApollo oil ATF-DX
Circulating lubrication: 100, 200, 300 ml/min.
Initial oil temperature: 80˚C

Criteria for
Detection of roller drilling wear※2,
stopping the test or occurrence of abnormal vibration
Run time

5 hours※2 Roller Drilling Wear: The outer end face of each roller is forced into the outside
diameter pocket face of the corresponding cage pocket owing to factors including
the centrifugal force as well as roller rotation on its own axis. As a result, wear canoccur on the cage pocket outside diameter face. This phenomenon is referred toas “roller drilling wear”.

Fig. 2 Diagrammatic illustration of NTN horizontal typethrust bearing test rig

Table 6 Externals situation of cage and roller edge side
Bearing temperature ˚C

200

Oil flow rate

180
160
140
120

Cage

100
80

100ml/min.

60

200ml/min.

300ml/min.

40
20
0

0

1

2

3

4

5

RollerTime hFig. 3 Temperature-rise comparisons of each oil
quantity-108-

100ml/min.

200ml/min.

300ml/min.

High Speed Thrust Needle Roller Bearings

min-1 (dmn = 86,000).
The allowable limiting speeds of the new design andconventional design at an oil flow rate of 100 mL/min
are graphically shown in Fig. 4.
The allowable limiting speed of the new design is atleast approximately 1.7 times higher then that of theconventional design (76% improvement over theconventional design).

(2) Test result
Under the test conditions in Table 7, the allowable
limiting speeds of the new design are as follows:
12,000 min-1 (dmn = 86,000) at oil flow rate of 100
mL/min; and 15000 min-1 (dmn = 1.07 million) at oil
flow rate of 200 mL/min.
In the temperature rise test, the bearing
temperatures of the new design and conventional
design were measured at the allowable limiting speedof the conventional design.
Fig. 5 illustrates the heat rise patterns for the newand conventional bearings at the allowable limiting
speed of the conventional bearing.
The new and conventional bearings were run at a
speed of 12,000 min-1 (dmn = 86,000) at an oil flowrate of 100 mL/min. The temperature on the new
design was approximately 24°C lower compared withthe conventional design. In other words, the new
design can limit the bearing temperature rise by 36%.
Also, the new and conventional bearings were run
at a speed of 15,000 min-1 (dmn = 1.07 million) at an oil
flow rate of 200 mL/min. The temperature on the new
design was approximately 25°C lower compared withthe conventional design. In other words, the new
design can limit the bearing temperature rise by 38%.
Compared with the conventional design, the newly
developed high-speed thrust needle roller bearing can
be characterized that the running temperature will be
low, and its heat generation is limited.

4. 4 Temperature rise test
For the temperature rise test, the NTN horizontaltype thrust bearing test rig shown in Fig. 2 was again
used.

(1) Test conditions
The test conditions for the temperature rise test aresummarized in Table 7.
dmn=1.51 million
New
designdmn=0.86 million
Conventional
design

0

5000

10000

Allowable limiting speedof the new design: morethan 1.7 times comparedwith the conventional
design.

15000

20000

25000

Allowable limiting speed min-1

Fig. 4 Comparison of Limiting speed
Table 7 Temperature-rise test condition
Test rig

NTN horizontal-type thrust bearing test rig

Speed

Allowable limiting speed with the conventional design

Axial load

2780 N

Lubrication
systemApollo oil ATF-DX
Circulating lubrication: 100，200 ml/min.
Initial oil temperature: 80˚C

Criteria for
stopping the test

At detection of occurrence of drilling wear,
or occurrence of abnormal vibration

Run time

5 hours

5. Conclusion
The NTN innovative high-speed thrust needle roller
bearing has been thus described.
This product will find a wide range of applications
including motors for hybrid vehicles, automatictransmissions and CVT’s on conventional vehicles, aswell as general industrial machinery.
NTN’s desire is to develop an optimized version of
the new high-speed thrust needle roller bearing foreach particular industrial field.

Bearing temperature ˚C

200

Conventional design

New design

180

Photo of author

160
140

Decrease of
24˚C

Decrease of
25˚C

Oil flow rate 100 mL/min
12000 min-1(dmn=86,000)

Oil flow rate 200 mL/min
15000 min-1(dmn=1.07 million)

120
100
80

Kosuke OBAYASHI

Fig. 5 Temperature-rise comparison between development
bearing and current bearing by limiting speed for current bearing

Needle Roller Bearing
Engineering Dept.
Automotive Sales Headquarters-109-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Article ]Accessory Technology Trends and Product Developments

Ikuo FUJINIWA*
Makoto MURAMATSU*
Tadahisa TANAKA*

In response to the demand for environmental compliance and highperformance, in the automotive industry, NTN has developed
accessory products that contribute to long-life, high performance,light weight, and compact design. In this paper we will introduce thedesign, features, and evaluation test results of these products.

the following chapters, the authors will present how
these NTN products are involved in, and contribute to,the evolution of automobiles.

1. Introduction
Automobiles have made a great deal of progressfrom the 20th to the 21st centuries with advances insafety, environment-friendliness (low fuelconsumption), comfort, and cost. These advances
have come about because factors such as the
progress from passive to active safety, enhanced
awareness of environmental protection, andrequirements for enhanced convenience are all
required from a single automobile.
At the same time, requirements for automotive
parts, represented by electrical and auxiliarie
components, have become more and more severe interms of elongation of life, improvement of reliability(safety), elimination of environment-burdening
substances, weight reduction, downsizing (increased
fuel economy), reduced vibration, enhanced
quietness, and the transition from manual to motordriven operation (comfort).
To respond to these requirements, NTN has
developed lightweight, compact electromagnetic
clutch bearings, alternator bearings responding toincreased electrical loading, auxiliary-component-use
auto-tensioners improving the quietness and life ofbelts, and pulley units with built-in clutches. It goes
without saying that these products do not containenvironment-burdening substances.
Fig. 1 shows a sketch of an engine with typicalapplications of the NTN products presented above. In

The pulley with a built-in clutch mentioned above
will be separately described in a product presentation
featuring "alternator-use pulleys with a built-in small
clutch."Auto-tensioner for
auxiliary-component-use belts
Alternator bearing

Electromagnetic-clutchuse bearing
Pulley with a built-in clutch

Fig. 1 Engine and accessory device*Automotive Engineering Dept. Automotive Sales Headquarters-110-Accessory Technology Trends and Product Developments

based thickening agent suitable for outer ring rotation.
Table 1 shows the characteristics of the existinggrease and the newly developed product.
In addition to these, the features of the developed
bearing are as follows:

2. Double-row bearing forelectromagnetic clutches
Electromagnetic clutches, for automotive airconditioners, are used in environments allowing
external moisture and dust to enter. For this reason, it
is essential for such bearings to be provided with highsealing performance in order to ensure the service lifeof the bearings. In addition, the recent requirement of
increased reliability makes the long life durability ofgrease under high temperature a problem that needs
to be solved. Being successful in developing bearings
responding to these requirements, we would like topresent them in this commentary.
The seal developed by NTN was arrived at with
FEM results and quality engineering; such as
optimized seal lip contact angles, end shapes, andsurface coarseness of the sliding surface, exhibiting
stable sealing characteristics from low to high speed
rotation. Fig. 2 shows the cross-section of a doublerow bearing for an electromagnetic clutch, and Fig. 3
the seal lip shape of an existing product and the newly
developed one. Regarding the elongation of grease
life, we have developed a long-life grease with a life
double that of the existing product by using a urea-(1) Improvement in the resistance to muddy water･ Suppression of the invasion of muddy water to about
1/10 of the existing bearing design(Fig. 4)･ Assurance of stable resistance to muddy water at
high-speed rotation･ Capable of accommodating thin bearings (with across-sectional height that is 70% of that of standard
products)

Table 1 Grease properties

Thickening agent
Base oilBase oil viscosity mm2/s（40˚C）
Rust development percentage %Sodium nitrite

Our existingproduct

Developed
product

DiureaDiureaEther oil＋PAO

Ether oil

72.3

100

10.2

0.8

Contained

Not containedElectromagnetic clutch

Our existingproduct

700

Testing conditions

0.35

Number of revolutions per minute : 2,000min-1
Load : 1,960N
Amount of muddy water : 1L/min 10wt%
Testing time : Three hours

0.3
0.25

600

0.2
0.15
0.1

500

Testing conditions
Number of revolutions per minute : 7,000min-1
Load : 1,470N
Offset : 13mm
Bearing temperature : 160˚C
Ambient temperature : 120˚C

400
300
200

0.05
0

Fig. 3 Cross-section of seal lip

800

0.4

Seizing life hQuantity of muddy water invasion g

Fig. 2 Cross-section of double-row bearing for electromagnetic clutch

Developed
product

100

Competitor’s
product

Our existingproduct

Developed
product

0Fig. 4 The evaluation test result of seal nature

Competitor’s
product

Our existingproduct

Developed
product

Fig. 5 The evaluation test result of durability-111-

NTN TECHNICAL REVIEW No.75（2007）

alternator bearings that are capable of being
continuously operated, even under a temperature of
180˚C, which we would like to present below.
Fig. 6 shows the structural sketch of an alternatorand Fig. 7 the schematic cross-section of the
bearings. Attached to the housing, the bearings
support shafts for the rotor, fan, and pulley.
Mechanical load resistance is required of the front
bearing that is close to a load point (the pulley), while
heat resistance is required of the rear bearing that isplaced in an environment with poor heat dissipation.
The features of component parts forming the bearings
are as follows:(2) Long life at elevated temperatures･ Assurance of a life about two times that ofconventional products (Fig. 5)(3) Environmental friendliness･ Grease does not containing sodium nitrite

3. Alternator-use high temperature,long-life ball bearings
Originally, direct current generators called dynamos
were used as automotive generators. However, an
increase in the placement of electrical components
has lead to extensive use of alternators that arecompact and yet capable of generating large power
outputs. Two deep-groove ball bearings are commonlyused on an alternator, contributing to reducing the size
and weight of an alternator and increasing its output.
Presently, various on-board components have been
made electronically operated to enhance comforts,
safety, and fuel economy. A further increase in power
generation of the alternator is then required. For thisreason, it is foreseen that the temperature of a bearing
in operation will rise from the current 150˚C to 180˚C.
Bearings are needed that last long and are free fromthe flaking peculiar to alternator bearings (brittle
flaking) under high temperatures. We have developed

Front bearing(1) Bearing materials
High carbon chrome bearing steel (SUJ2) is usedas the bearing material to achieve the goal of
economy and long life. In addition, special heat
treatment is applied to the material to secure
dimensional stability at high temperature and to
improve brittle flaking resistance. Fig. 8 shows theresults of brittle flaking life tests.

(2) CageWith loading due to lead and lag of the balls
constantly applied to the cage, it need to be highstrength. In addition, holding a ball in a stable

Rear bearing

Front bearing

Fig. 6 Cross-section of alternator

140

Time

h

120
100
80

Fig. 7 Cross-section of bearing and seal

Bearing : 6303
Number of revolutions per minute: 0〜20,000min-1
Radial load : 1,617N
Current : 1.0A
Ambient temperature: Room temperature

10000
ACM
Newly developed material

Time h160

Rear bearing

60

1000

100
40
20
0

Newly
developed bearing

10

SUJ2-madestandard bearing

150

160

180

Temperature ˚CFig. 8 Test result of brittle flaking life

Fig. 9 Durability by seal material

-112-

200

Accessory Technology Trends and Product Developments

condition requires the pocket gap to be stable in a wide
temperature range. Made of aromatic polyamide resinmaterial excellent in high temperature characteristics,
the cage responds to these requirements.

often used as housing material. If the creeping of abearing while driving causes wear in the housing, thepower generating performance is adversely affected.
To prevent housings from being worn, NTN has
developed expansion-compensation bearings with aresin ring placed on the outside diameter of the outerring. This ring has a large linear expansion coefficientand excellent heat resistance.

(3) Seal
Newly developed heat-resistant acrylic rubber isused as the rubber material to make the seal operable
at 180˚C. Fig. 9 shows the heat-resistant life curve of
the seal rubber material.
In addition, the seal tolerance is specified to always
have interference contact; this allows a lip shape to beformed that requires low torque and prevents greaseleakage and intrusion of dust.

4. Auto-tensioners for auxiliarycomponent belts

A urea-based thickening agent with reinforcedhydrogen bonding strength is used. A newly invented
additive helps increase high temperature life by 50%
when compared with conventional products. Table 2
shows the characteristics of the newly developed
grease and Fig. 10 the result of the life test for thegrease capable of being used at 180˚C. At the sametime, the brittle flaking life has been increased by twotimes that of conventional products.
In addition, this grease is environment-friendly
grease and does not contain sodium nitrite.

In response to a recent trend of environmental
friendliness, the serpentine scheme in which different
auxiliary components are driven by a single belt to
reduce engine size and weight is often adopted. Thisrequires auto-tensioners to adjust belt elongation andtension.
Auto-tensioners for auxiliary component belts areclassified into the friction type and the hydraulic typein terms of damper structures. NTN adopts thehydraulic type for the reasons of high damper force
and high reliability.
The features of NTN auto-tensioners are described
below, with the structure shown in Fig. 11.

(5) Expansion-compensation bearing(1) Lightweight

In the rear bearing, a loose fit is usually used
between the outer ring and the housing. Aluminum is

The cylinder, forming the casing for the hydraulicdamper section, and the spring washers are made of
die-cast aluminum, and they are combined with steel
valve sleeves to achieve reduction in weight and wear
resistance.

(4) Grease

Table 2 Grease properties
Newly developed
greaseConventional
greaseThickening agent

Aromatic urea

Aromatic urea

Base oil

Ether+PAO

Ether+PAO

53.5

72.3

8.5

10.1

Mixed consistency（60w）

286

300

Dropping point（˚C）

>260

227

40˚C
Base oil dynamic
viscosity mm2/s 100˚C(2) High reliability
To seal the hydraulic fluid in the hydraulic damper
section, an oil seal is adopted. The damper is formed
in a gas-liquid two-layer structure to ensure thereliability against oil leakage.
High-strength rubber bushings are employed on theconnecting section on both pulley arm and engine side
to enhance wear resistance and usability under hostileenvironments.

(3) Tuning
1000

In common reciprocating engines, auto-tensioners
are subjected to sinusoidal load arising from the change
in angular velocity of the engine crank (Fig. 12).
In addition, as shown in Fig. 13, variations in belttension occur in the engine revolution range
depending on the load on the engine and auxiliarycomponents. Since large change in the belt tension
occurs in the low revolution range (2000 min-1 orlower), high damper force is often required.
The attenuation of a hydraulic damper is a onedirection damper as shown in Fig. 14. The damper
force can be set at an optimum value such as a highdamper or low damper specification depending on the

Test temperature

Time h800: 180˚C
Number of revolutions per minute: 10,000min-1

600

400

200

0

Newly developed
greaseConventional
greaseFig. 10 Test result of grease life-113-

NTN TECHNICAL REVIEW No.75（2007）

Connecting section (rubber bush)
Pulley section

Spring washer (aluminum die-cast)
Cylinder(aluminum die-cast)

Retaining ring
Oil seal
Hydraulicdamper section

Air

Rod

Wearing

Return spring

Oil
PlungerCheck ball

Valve sleeve (steel)

Valve retainer
Plunger spring
Arm section
Connecting section (rubber bush)Fig. 11 Sectional view of auto-tensioner for accessory belt

3000

Maximum load

2000
1000
Minimum load

0
360

0

720

1080

1440

Crank angle deg

Fig. 12 Angular velocity of crank pulley

Load occurring in the tensioner N

Tensioner input load N

5000
4000

High damper specification
4000

3000
Low damper
specification
2000

1000
Spring force

Tensioner input load N

0
-0.6

4000

"Large change
in the belt tension"

-0.4

-0.2

0.0

0.2

Amplitude

mm

0.4

0.6

Fig. 14 Damping characteristics

3000
2000

Maximum load

1000
Minimum load

0
0

1000

2000

3000

4000

5000

6000

7000

Engine number of revolutions min-1

Fig. 13 Auto-tensioner input load variation depend on engine speed

Fig. 15 Compact auto-tensioner for accessory belt-114-

Accessory Technology Trends and Product Developmentscharacteristic of each engine. In addition, the return
spring is arranged outside the case to be capable ofaccommodating various belt tension requirements.
Furthermore, we will develop auto-tensioners forsmall auxiliary-component belts (Fig. 15) seeking afurther reduction in size, weight and cost.

5. Concluding remarks
In this commentary, we have described the recent
trend in automotive electrical components andauxiliaries, along with NTN’s approaches to thesetrends. We hope that these electrical components andauxiliaries will be spread in the market as products
responding to requirements for automotive safety andenvironmental -friendliness.
Anticipating that requirements for automotive
electrical components and auxiliaries will become
increasingly sophisticated, we will push forward withthe development of high-performance products
capable of addressing required environmental
conditions.

References
1) Satoshi Kitano, Tadahisa Tanaka, Tomokazu
Nakagawa: NTN Technical Review, No.73 p.116(2005)

Photos of authors

Ikuo FUJINIWA

Makoto MURAMATSU

Tadahisa TANAKA

Automotive Engineering Dept.
Automotive Sales Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters-115-

NTN TECHNICAL REVIEW No.75（2007）[ Technical Paper ]Development of NA103A Long-life Grease
for Automotive Components

Takayuki KAWAMURA*
Hidenobu MIKAMI*
Recently, as automotive components have become progressively
more compact, ball bearings are required to operate at higher
temperatures, rotational speeds and loads. In such cases, grease lifeand brittle flaking life have become more critical for determining thebearing’s overall life than rolling contact fatigue life. This reportpresents a novel and unique method to generate brittle flaking in ashort time, and also introduces the newly developed NA103A grease
for automotive electrical instruments and auxiliary devices. NA103A isa great improvement over common greases, including enhanced
endurance under high temperatures and anti-brittle flaking performance. Moreover, this grease does not containhazardous materials such as sodium nitrite and barium compounds. NA103A is expected to be useful in several
types of automotive applications.

and auxiliaries,distinct premature flaking (brittle
flaking) occurred, which was accompanied by white
structural change on the rolling surface of a ballbearing, posing a problem 4) 5),. Although the adoptionof grease consisting of urea-ether base oil as acountermeasure has substantially reduced cases ofnonconformance in the market 6), the trend toward
increasingly severe working conditions threatens topose this problem of brittle flaking again. However,since brittle flaking is difficult to reproduce in a shorttime, sufficient experimental verification has not beenaccomplished, thus hindering the possibility of taking
effective counter measures.
This paper presents a unique method ofreproducing brittle flaking in a short time that has been
worked out from the viewpoint of friction. This paper
also presents NA103A grease for electricalcomponents and auxiliaries developed on the basis of
the knowledge obtained through evaluation of thismethod. This grease has excellent durability at hightemperatures, resistance to brittle flaking, cold hootnoise suppression, and anti-rust resistance.

1. Introduction
To respond to recent requirements for the reduction
in size, weight, and noise of automobiles, electricalcomponents and auxiliaries (alternators,
electromagnetic clutches, auxiliary component pulleys,
and the like) are being downsized and slimmed. On
the other hand, these devices require increased
output and efficiency. As for electrical components
and auxiliaries, the decrease in output resulting fromdownsizing must be overcome by increasing therotational speed. For this reason, ball bearings tend to
be increasingly exposed to high temperatures.
Notably, bearings for alternators have required more
and more power and it is estimated that the working
temperature will be raised from the current 150˚C to alevel of 180˚C. Since the life of grease is thought tobe shortened substantially under high temperature
conditions 1-3), it is expected that conventional grease
will not meet future service life requirements.
In the period from the latter half of the 80s to theformer half of the 90s, V-belts were replaced by Vribbed belts as driving belts for electrical components*Elemental Technological R&D Center-116-Development of NA103A Long-life Grease for Automotive Components

verification has been difficult to perform.
The authors conducted sliding tests in oil using aball-on-disk testing device (Fig. 2) found the tests
accompanied by heavy wear were coupled with alarge amount of hydrogen intrusion into the steel (Fig.
3) 11-13). Paying attention to the fact that wear
accelerates intrusion of hydrogen, we then operated
the bearing with current passing through the bearing
to accelerate wear and thereby reproduced brittle
flaking in a short time 14-16).

2. Method of reproducing brittle flaking
in a short time
In bearings for automotive electrical components
and auxiliaries, early stage flaking (brittle flaking)accompanied by white structural change occurs insome rare cases. NTN thinks that hydrogen
embrittlement flaking is generated by the mechanism
shown in Fig. 14-10). Although a number ofcountermeasures were taken, experimental

Slip of a rolling body caused by vibration and abrupt
acceleration or deceleration
Rolling body
Lubricant

Slip

Generation of a new steel surface causedby metallic contact
Decomposition of the lubricant causedby tribochemical reaction

Rolling surface

Hydrogen embrittlement caused by the generation ofand intrusion into steel by hydrogen

Brittle flaking
New steel surface

Total hydrogen content in steel ppm

Fig. 1 Estimated mechanism for brittle flaking

Test condition
Steel ball : Made of SUJ2
Ring : Made of SUJ2（φ24.1mm×φ17.3mm×1mm）
Oil temperature : 100˚C
Slip speed : 0.3m/s
Bearing stress : 3.0GPa

Load

Load
Steel ball

Load

Oil bath

1.5

1

0.5

0
0

5

10

15

Ring wear depth μm

Ring test
piece
HydrogenanalysisFig. 2 Test apparatus for sliding test in oil

Before the experiment

Water glycol

Mineral oil

PAO

Ester

Ether

Fig. 3 Relationship between amount of hydrogen in steel
and wear depth-117-

NTN TECHNICAL REVIEW No.75（2007）

2. 1 Experiment method

2. 2 Results of the experiment and theirconsideration

Fig. 4 shows a diagrammatic sketch of the testingmachine. The specimen bearing and the support
bearing support the rotation axis of this testingmachine. With the housing divided into left and righthalves with an insulating material inserted between
them, the testing machine is structured so it allows
current to flow from terminal A through the specimen
bearing to the rotation axis, then through the support
bearing, and finally to terminal B when electricity isapplied between terminals A and B.
Tables 1 and 2 shows the characteristics ofspecimen greases and the experiment conditions. The
experiment was conducted with a constant load
current applied under the condition of abrupt
acceleration and deceleration of a rotating inner ring.
In another experiment, the direction of the current was
reversed to study the effect of the reversing thecurrent.

Terminal A

Table 3 shows the results of the experiment. Withno current flowing, both grease A and B allowed
operation to continue over 300 hours. Under theelectrically charged condition, both types of grease
allowed flaking to occur in a short time. Flaking
occurred in all of the inner rings, outer rings, and steel
balls. Fig. 5 depicts where the flaking occurred alongwith a cross-sectional photo. The cross-section shows
white structural change that is characteristic of brittle
flaking.
Fig. 6 shows the relationship between current andbrittle flaking life. Under electrically charged condition,
the larger the magnitude of current, the earlier flaking
occurs. When a current of 3-A flows, the flaking life isin the order of 10 hours. A current magnitude of 3-A
does not reveal the difference in life between grease A
and grease B. However, with smaller currents, theservice life differs depending on the type of grease.
With grease B, for which emphasis is placed has
greater resistance against brittle flaking than grease A
thus exhibiting a longer life.
As shown in Fig. 6, the bearing sealed with grease
A allows flaking accompanied by white structuralchange to occur after having been operated for about
20 hours under an electrically charged condition with acurrent of 0.5 A. To study the correlation of flaking
with hydrogen embrittlement, a bearing sealed withgrease A was operated for 3 to 15 hours with a current

Terminal B

Specimen bearing

Support bearing

Insulating
material
Pulley(belt-driven)

Fig. 4 Bearing test equipment

Table 3 Test results
Cur- Time to
No. Grease rent flaking
AhTable 1 Test greases
GreaseGrease A

Thickening agent
Base oil

Grease B

Urea

Urea

1

PAO＋Ester

ADE＋PAO

2

Dynamic viscosity mm2/s (40˚C)

40

72

3

Consistency

250

300

4
5

Specimen bearing

Deep-groove ball bearing (6203)

Amount of enclosed grease (g)0.86
Abrupt acceleration or deceleration
between 0 and 20000

Rotational speed (min-1)

0.1

88.0

―

Steel ball

20.7

―

Steel ball

19.0

Steel ball

―

22.1

―

Steel ball

8.6―

Outer ring, Steel ball

12.4

Inner ring

―

13.1

Inner ring

―

3.0

13.6

Inner ring, Steel ball

―

0

＞300

―

―

11

0.1

＞150

―

―

12

0.3

81.0

―

Steel ball

32.3

―

Steel ball

45.8

―

Outer ring

16.8

―

Steel ball

9.5―

Outer ring

9

Pulley load (N)

1,617

13

Bearing load (N)

2,332

14

0.1〜3
When vibration has become 10 times
that at the beginning

15
16

-118-

―

10

Room temperature

Stopping condition

―

2.0

8

Atmosphere

Current (A)

＞300

0.5

7

Table 2 Test conditions

Inner ring to outer ring Outer ring to inner ring0

GreaseA

6

Flaked portion
Direction of current

Grease
B

0.5
1.0
2.0

Development of NA103A Long-life Grease for Automotive Components

high probability be due to hydrogen embrittlement.
As shown in Table 3, flaking occurs in different
portions depending on the direction of the current.
When the current flows from the inner to the outer
ring, flaking occurs on the inner ring or steel balls,
while the current in the opposite direction causes
flaking to occur on the outer ring or steel balls. In otherwords, flaking occurs only on the positive side. To
explain the reason why flaking occurs on the positiveside only, we ran a thrust ball bearing alternating thedirection of the current; from this test, it turned out that
portions on the positive side had been selectively
worn 15).
For the above reason, as shown in Fig. 8, allowing
current to flow results in accelerated wearing on thepositive side of the bearing, with the new surface
produced by the wearing in turn causing grease to
decompose with the generation of hydrogen. It isestimated that the intrusion of the hydrogen generated
in this way leads to brittle flaking.

of 0.5 A. In addition, the quantity of hydrogen in thesteel balls were measured 30 minutes after the
stopping the test. Using a high temperature gas
chromatograph apparatus, the quantity of hydrogen
released while being heated up to 400˚C at a rate of
180˚C/h was used as measurement. Fig. 7 shows therelationship between the running time and the quantityof hydrogen. The longer the operating time, thegreater the quantity of hydrogen in steel becomes.
From the results described above, the flaking in a shorttime observed in this experiment can be said with a

Flaking on a raceway

Time to bearing flaking h200

Grease A
Grease B

150

100

50

0
0.01

0.1

1

10

Electric current A

Cross-sectional view of tested bearing
Fig. 5 Flaking (outer ring)

Fig. 6 Relationship between electric current and bearing life

0.30

Hydrogen content in steel ppm

Hydrogen release rate ppm/min

0.008

15h

0.006

10h
0.004

3h

0.002

0.000
100

200

300

0.20

0.10

0.00

400

Temperature ˚C

0

5

10

15

Running time hFig. 7 Relationship between amount of hydrogen and running time

-119-

20

NTN TECHNICAL REVIEW No.75（2007）

Slip of a rolling body caused by vibration and abrupt
acceleration or deceleration

Current flows in the bearing

Generation of a new steel surface causedby metallic contact

Wear on the bearing rolling surface

Decomposition of the lubricant causedby tribochemical reaction
Rolling body

LubricantHydrogen embrittlement caused by the generation ofand intrusion into steel by hydrogen

Brittle flaking

Rolling surface

Fig. 8 Estimated mechanism for brittle flaking of bearing endurance test with electricalmixture of ether and PAO oil as the base oil. Provided
with adequate dynamic viscosity, the new grease canbe used in a wide range of temperatures including low
to ultra-high temperatures by suppressing the selfheating of the grease at high speed rotation.
What distinguishes the new grease fromconventional grease B is the passivation of newly
produced steel surfaces with multiple additives toprevent brittle flaking and an improvement of hightemperature endurance. This grease does not containenvironment-burdening substances represented bysodium nitrite and barium compounds at all. Table 4shows the properties of grease NA103A.

3. Electrical component and auxiliary
bearing grease "NA103A"High temperature durability, resistance to brittle
flaking, cold hoot-noise suppression, anti-rustresistance, and similar characteristics are required ofgrease for electrical components and auxiliaries. As
shown in Table 4, all these type of greases consist of
urea grease using synthetic oil as base oil; greaseaiming at cold hoot noise suppression (conventional
grease A) and grease aiming at resistance to brittle
flaking (conventional grease B) are used. To replacethese kinds of conventional grease, we havedeveloped a new type of grease, NA103A that
improves all characteristics required for use withbearings for electrical components and auxiliaries.

3. 2 High temperature resistanceGrease NA103A underwent a high temperature
endurance test (according to ASTM D3336) at 150˚C
and 180˚C.
Fig. 9 compares the results obtained fromconventional grease and NA103A. Grease NA103A
exhibits durability 1.5 to 2 times that of conventional

3. 1 Composition and properties of NA103A
greaseLike conventional grease B, grease NA103A uses
urea compounds as the thickening agent and a

Table 4 Greases for automotive components
NA103A

Conventional
grease A

Conventional
grease B

Urea

Urea

Urea

Ether＋PAO

PAO＋Ester

Ether＋PAO

54

40

72

JIS K2220.23

Worked penetration 60W 25˚C

286

250

300

JIS K2220. 7

Dropping point ˚C

260＜

250

240

JIS K2220. 8

Oil separation mass% 100˚C 24h

3.0

0.2

1.7

JIS K2220.11

Oxidation stability kPa 99˚C 100h

20

20

20

JIS K2220.12

Start

271

220

270

Rotation

63

45

78

GreaseThickening agent
Base oil
Dynamic viscosity mm2/s (40˚C)

Low temperature torque mN・m -30˚C-120-

Test method

JIS K2220.18

Development of NA103A Long-life Grease for Automotive Components

Test condition
Bearing: Bearing 6204

Load: Fa＝Fr＝67N

1000
Grease life（180˚C）h

5000
Grease life（150˚C）hRotational speed: 10000min-1

Amount of grease sealed: 1.8g

4000
3000
2000
1000

800
600
400
200

0

0
NA103A

Conventional
grease A

Conventional
grease B

NA103A

Conventional
grease A

Conventional
grease B

Fig. 9 Bearing endurance test results under high temperature

grease A and B under both of these temperature
conditions. With an addition of optimum anti-oxidants,
grease NA103A inhibits oxidation-based deterioration,
exhibiting long life under high temperature conditions.

Test condition
Bearing: Bearing 6203
Amount of grease sealed: 0.86g
Rotational speed: Abrupt acceleration or
deceleration between 0 and 20000 min-1
Load: Fr＝2332N
Current: 0.5A

3. 3 Resistance to brittle flaking
60

Brittle flaking life h

We evaluated the resistance to brittle flaking ofgrease NA103A by means of the current passage
acceleration test method described in section 2.
As shown in Fig. 10, NA103A has resistance tobrittle flaking more than twice that of conventional
grease A and 1.5 times that of conventional grease B
thanks to the effect of the additives passivating new
surfaces on the steel.

40

20

0

NA103A

3.4 Cold hoot-noise suppression characteristic
In belt pulleys equipped with deep-groove ballbearings, cold abnormal noise called hoot noiseoccurs occasionally at the time of starting in cold
regions depending on the pulley specification andoperating conditions. This abnormal noise has afrequency range between 500 and 1000 Hz, with theacoustic pressure exceeding 100dBA occasionallyand thereby producing very grating noise. The cold
abnormal noise is believed to be generated by theaxial vibration of the outer ring (so-called translation
motion) resulting from the resonance of the pulleysystem caused by the self-excited vibration of therolling body. It is known that the type of grease affects
the generation of noise greatly.
A specimen bearing incorporated into the test
pulley, which is cooled at -60˚C for five hours, isfastened to the testing machine shown in Fig. 11. The
testing machine is started when the bearing
temperature reaches -20˚C, and is kept operating until
the bearing temperature reaches +10˚C withtensioning on the belt. The rate of occurrence of cold

Conventional
grease A

Conventional
grease B

Fig. 10 Brittle flaking test results

Idler pulley
Test pulley

Drive pulley

Timing belt

Fig. 11 Hoot noise measurement-121-

NTN TECHNICAL REVIEW No.73（2005）

hoot phenomena during this period was confirmed. As
shown in Fig. 12, grease NA103A has a cold hootsuppression characteristic comparable to that ofconventional product A that is excellent in low
temperature properties.

4. Concluding remarks
Compared with conventional products, grease
NA103A excels in all of high temperature durability,
resistance to brittle flaking, cold hoot suppression
characteristic, and anti-rust resistance, which arerequired of grease used with bearings for automotive
electrical components and auxiliaries. In particular,
provided with high temperature durability andresistance to brittle flaking, which are more than 1.5
times those of conventional products, the new greaseis expected to be applied to bearings for alternators ofwhich higher power generating capability is required
and that are exposed to a high temperatureenvironment.
In addition, composed of safer additives, base oil,and thickening agent, the new grease is expected togreatly contribute to the enhancement of performanceof automotive electrical components and auxiliaries
and the reduction in environment loading alike.

3. 5 Anti-rust resistance

Rate of occurrence of cold hoot %Fig. 13 shows the result of an anti-rust resistancetest on NA103A. To evaluate NA103A, we calculated
the rate of occurrence of rust on a tapered roller
bearing (30204). The bearing was coated with thegrease under the test, immersed in a 1%-sodiumchloride aqueous solution and left alone in a 40˚C,moist atmosphere for 48 hours. Then the outer ring's
rolling surface was divided circumferentially into 32
equal sections. Then the number of sections that
contained rust was divided by 32. We confirmed,
thereby, that grease NA103A has anti-rust resistanceequal or superior to that of conventional products.

References
1001）Booser, E. R. (1974), "Grease Life Forecast for Ball
Bearings", Lubrication Engineering, 30, pp 536-541.
2）Kawamura, T. Minami, M. and Hirata, M. (2001),"Grease Life Prediction for Sealed Ball Bearings",
Tribology Transactions, 44, 2, pp 256-262.
3) Kawamura, T., Minami, M., Hirata, M. (2001).
Projection of Life of Sealed Ball Bearing Grease, NTN
TECHNICAL REVIEW No.69, pp76-81.
4) Tamata, K., Maeda, K., Tsushima, N. (1992). New
Type of Micro-structural Change Observed in Bearings
for Electrical Components and Auxiliaries, NTN
TECHNICAL REVIEW No.61, pp29-35.
5) Nozaki, S., Fujimoto, K., Tamata, K. (1992). Brittle
Flaking in Bearings for Electrical Components and
Auxiliaries and Extension of the Life of Bearings, NTN
TECHNICAL REVIEW No.61, pp36-39.
6) Nozaki, S., Okasaka, M., Kubota, Y., Akabe, S. (1996).
Technical Trend in Bearings for Electrical Components
and Auxiliaries, NTN TECHNICAL REVIEW No.65,
pp65-72.
7) Tamada, K. and Tanaka, H. (1996), "Occurrence ofBrittle Flaking on Bearings used for Automotive
Electrical Instruments and Auxiliary Devices," Wear,
199, pp 245-252.
8) Kino, N. and Otani, K. (2003), "The Influence ofHydrogen on Rolling Contact Fatigue Life and its
Improvement, " JSAE Review, 24, pp 289-294.
9) Kino, N., Yamamoto, K., Otani, K., Uchiyama, N.
(2004). Analysis of Factors of Hydrogen Embrittlement
Type Early Flaking, Proceedings of the Japan Societyof Mechanical Engineers (Edition A), 70, 696, pp54-61.

75

50

25

0

NA103A

Conventional
grease A

Conventional
grease B

Rust development percentage %Fig. 12 Rate of hoot noise generation at low temperature

40

30

20

10

0

NA103A

Conventional
grease A

Conventional
grease B

Fig. 13 Rust prevention ability of greases-122-Development of NA103A Long-life Grease for Automotive Components

10) Hamada, H., Matsubara, Y. (2006). Effects ofHydrogen on Tensile and Compression Fatigue andRolling Fatigue of Bearing Steel, NTN TECHNICALREVIEW No.74, pp50-57.
11) Kawamura, T., Ohara, M., Tamata, K. (2004).
Measurement of the Amount of Intrusion of Hydrogen
into Steel in a Ball-on-disk Test in Oil, Abstract for
Tribology Conference, (Tokyo, May 2004), pp173-174.
12) Kawamura, T., Ohara, M., Tamata, K. (2004).
Measurement of the Amount of Intrusion of Hydrogen
into Steel in a Ball-on-disk Test in Oil (Second Report),
Abstract for Tribology Conference, (Tottori, November2004), pp593-594.
13) Kohara, M. Kawamura, T. and Egami, M. (2006),"Study on Mechanism of Hydrogen Generation from
Lubricants", Tribology Transactions, 49, pp 53-60.
14）Mikami, H. and Kawamura, T. (2007), "Influence ofElectrical Current on Bering Flaking Life", SAE
International, 2007-01-0113.
15) Kawamura, T., Mikami, H. (2007), Effect of Current
Passage on Bearing Flaking, Abstract for Tribology
Conference, (Tokyo, May 2007), pp267-268.
16) Kawamura, T., Mikami, H. (2007). Effects of Grease
Additives on Brittle Flaking under Energized Condition,
Abstract for Tribology Conference, (Saga, September2007), pp57-58.

Photos of authors

Takayuki KAWAMURA

Hidenobu MIKAMI

Elemental Technological
R&D Center

Elemental Technological
R&D Center-123-

NTN TECHNICAL REVIEW No.75（2007）[ New Product ]Compact Clutch Integrated Pulley for Alternators

Koji SATO*
Isao MIKURIYA*

The demand for a Clutch Integrated Pulley (CIP) for alternators is increasing,as automakers try to reduce belt slip noise and fluctuation especially forengines with large torque vibration. NTN has developed a compact CIP with thesame torque capacity as our standard type, making it possible to rotate at
higher speeds.

1. Introduction

2. Part to which a CIP is fastened

Fuel-efficient diesel engines work on highcompression ratios, which results in large fluctuationsof engine torque. A clutch integrated pulley(abbreviated below as CIP) is mounted on thealternator of a diesel engine that has large torquefluctuation. The reduction in belt tension fluctuation
improves the belt service life and prevents belt slip
noise from occurring. In gasoline engines the adoptionof direct injection fuel systems and the lowering of
idling rotational speeds for better fuel efficiency tends
to lead to an increased fluctuation in belt tension.
These circumstances make the usage of CIP’s anecessity. There is also a trend for reduced vehicle
weight and downsized engine peripherals, whichstrengthens the need for downsizing and slimming
engine auxiliaries. For diesel engines characterized bylarge torque fluctuation, high torque capacity isneeded; while for gasoline engines that operate athigh rotational speeds, responsiveness at high-speed
rotation is required.
NTN has developed a small-sized CIP withimproved high-speed idling performance, while still
maintaining the same high torque capacity as theconventional product. This was accomplished bydecreasing the shaft diameter, optimizing the designof the sprag clutch and support bearing, and adopting
high-performance grease. This report presents thissmall-sized CIP.

Fig. 1 shows the accessory component drive belt
system to which the CIP is fastened. A belt conveying
power from the crank pulley of the engine drives
accessory components such as an alternator, acompressor for the air-conditioner, and a power
steering pump. Among these components, thealternator has the largest moment of inertia. For thisreason the CIP is fastened to the alternator in order toalleviate fluctuation in belt tension.

Crank pulley

CIP

AlternatorFig. 1 Front end accessory drive system*Automotive Engineering Dept. Automotive Sales Headquarters-124-

Compact Clutch Integrated Pulley for Alternators

sprag is turned clockwise both by the outer ring andby the shaft the height of the sprag becomes higher
than the height of the sprag space. This causes theouter ring and the shaft to come into engagement, asshown in Fig. 4. When the outer ring turns slower than
the shaft, that is, when the sprag is turned
counterclockwise both by the outer ring and by theshaft the height of the sprag becomes lower than theheight of the sprag space. This causes the outer ringand the shaft to turn without engagement.

3. Structure and Functions of a CIP
3. 1 Structure of a CIP
Fig. 2 shows the cross section of a CIP. A spragclutch is placed between the shaft and the pulley withthe sides of the clutch supported by two bearings.
The clutch section consists of a cage with sprags
and springs inserted into pockets of the cage, forming
a one-way clutch between the outer ring that is pressfitted into the pulley and the shaft.

Outer ring

PulleyOuter ring

Support bearing

SpringDo

Outer ring

Ds

Cage
Shaft

Clutch in engagement

Clutch out of engagement

Shaft

Fig. 4 Operational principle of sprag clutch
Shaft

3. 3 Functions of a CIP
When an engine that is greatly fluctuating inrotational speed drives an auxiliary component with alarge moment of inertia, such as an alternator,
fluctuation in torque occurs in the pulley and belt drive
system. For this reason the belt undergoes large
fluctuations in tension that can shorten its life. To solve
this problem a one-way clutch is incorporated into thepulley to absorb the fluctuation in rotation caused by
the fluctuation in torque. This reduces the fluctuation intension acting on the belt, which results in increased
belt life.
As shown in Fig. 5, when the engine speed (V1)
decreases due to fluctuation in rotation, the alternator,
which has a large moment of inertia, acts to maintainits speed (V3). When V2 < V3 the clutch runs without
engagement to absorb the fluctuation in rotation; this inturn lowers fluctuation in belt tension, thus contributing
to suppressing noise caused by the slipping belt whenthe engine is accelerated or decelerated harshly.

Sprag

Fig. 2 CIP structure

3. 2 Operating Principle of a Sprag Clutch
As shown in Fig. 3, a sprag is designed to change
its height as its inclination changes. For example,
when a sprag rolls counterclockwise, the height of the
sprag decreases (H1) and when it rolls clockwise, theheight increases (H2), with the relationship between H1and H2 expressed as H1 < H2.
In contrast, the space formed between the inside
diameter of the outer ring into which the sprags areinserted (Do) and the outside diameter of the shaft (Ds)is constant (shown in Fig. 4). This is referred to in thefollowing as the sprag space, (Do-Ds)/2. When theouter ring turns faster than the shaft, that is, when a

V2 : Rotational speedof the pulley

H2

H1

V1 : Rotational speedof the engine

V3 : Rotational speed of the alternatorFig. 3 Sprag Tilting

Fig. 5 CIP Operation-125-

NTN TECHNICAL REVIEW No.75（2007）

Fig. 6 data on fluctuation in rotational speed and inradial load for both a fixed pulley and thecorresponding CIP. In this measurement the radialload acting on the pulley was measured as a
substitute characteristic for belt tension. It can be seenfrom the figure that large fluctuations occur in radialload (belt tension) of the fixed pulley while thefluctuation in radial load is substantially reduced for
the CIP.

2200

1800

1500

1600

1000

1400

500

1200
1000
0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10

Time

sec
Alternator shaft rotational speed
Pulley rotational speed
Radial load

CIP adopted
2500

2200
2000

2000

1800

1500

1600

1000

1400

500

1200

0
0.00

These efforts have provided the developed productswith the following features:･ High torque capacity: equivalent to that of aconventional product (rated torque: 90 Nm)･ Reduced size: reduction in the pulley diameterby 7 mm compared with that of a conventional
product (51-mm diameter min.)･Weight reduction: approximately 20% reduction
in weight･ Improved high-speed idling performance1000
0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10

Time

sec

（φ42）

（φ58）

（φ35）

（φ49）

Fig. 6 Comparison of Idling test between solid pulley
and CIP（a）Conventional product

（φ51）

2000

3000

Rotational speed min-1

2000

（φ25）

Alternator shaft/Pulley rotational speed
Radial load

Radial load N

Fixed pulley

2500

0
0.00

We have successfully reduced pulley diameters
without adversely affecting their high torque capacities
for our conventional products, which have been in
production since 2000 and are primarily used withdiesel engines. In addition, we have developed CIP’s
capable of being used under the high rotational speed
conditions of gasoline engines. Fig. 7 compares thedeveloped compact and conventional products. The
developed product incorporates an optimized design
specification on the following points.
(1) Reduction in the pulley diameter through the
reduction in the shaft diameter and support bearing
diameter(2) Securing of high torque capacities through the
improvement of sprags (the enlargement of thecross section and curvature)(3) Improvement of high-speed idling performancethrough the adoption of high-performance grease(4) Improvement of ease of manufacturing through thestraightening of the shaft

Radial load N

Rotational speed min-1

3000

4. Features of a Small-sized CIP（b）Developed product

Item

Conventional productDeveloped product

Shaft

37-mm outside diameter, with steps

25-mm outside diameter, without steps

Pulley diameter

58-mm diameter min.

51-mm diameter min.

Support bearing

6807 with a special outer ring

6905 standardSprag

4 mm in height×8.8 mm in length, 16 pieces

6 mm in height×12.2 mm in length, 10 pieces

Spring

16 pieces

10 pieces

Weight520g

400g

Fig. 7 Specifications of current CIP and compact CIP-126-

Compact Clutch Integrated Pulley for Alternators

5. 2 Test for Selecting Grease

5. Evaluation Test

The selection of grease is important in order toimprove high-speed idling performance. Aftercomparing different types of greases, grease A wasselected that is excellent in thermal stability and oilfilm forming capability and hence contributes to theimprovement of wear resistance. Fig. 10 shows theresults of the evaluation based on the comparison ofgrease A and grease B.

Various single-unit tests were conducted to confirm
functions of small-sized CIPs in a belt drive system.
The following are results of three of the tests
performed.

5. 1 Functional Evaluation
5. 1. 1 Harsh Acceleration and Deceleration Test
In this test harsh acceleration and deceleration
between low and high engine rotational speeds is
reproduced to evaluate the idling characteristics of aclutch.
As shown in Fig. 8, this test allows one to confirm
that a clutch has an adequate idling function underharsh acceleration and deceleration conditions.

Grease A

Grease B
0

Rotational speed min-1

200

Lifetime

25000

Pulley rotational speed
Alternator shaft rotational speed

Idling domain

300

400

500

h

Fig. 10 Grease life comparison

20000
15000

5. 3 Evaluation of Durability
To evaluate the durability of a CIP, tests wereconducted for the following items to confirm sufficientdurability.
(1) Harsh acceleration and deceleration test: between
5000 and 20000 min-1(2) Rotation fluctuation test: 2500 min-1 15%(3) High-speed durability test: 18000 min-1

10000
5000

Engagement domain
0

0

5

10

15

20

Time

25

30

35

40

sec

Fig. 8 Results of Cycled speed test

6. Summary

5. 1. 2 Rotation Fluctuation Test

This report presented NTN’s high-performance,
small-sized clutch integrated pulley (CIP) that wasmade possible through a reduction in shaft diameter,
improvement in sprags, and adoption of highperformance grease. Future market development for
the CIP product will focus on diesel engines and low
fuel consumption gasoline engine applications, where
the addition of a CIP will enhance the reliability ofauxiliary component drive belts.

In a rotation fluctuation test, a condition in whichhigh frequency, large fluctuation in rotation (as in
idling) is reproduced to evaluate clutch engagement
characteristics.
As shown in Fig. 9, the test allows one to make surethat the engagement of the clutch poses no problems
even under high frequency fluctuation in rotation.

Engagement domain

Idling domain

4000

2000

3500

1900

3000

1800

2500

1700

2000

1600

1500

1500

1000

Pulley rotational speed
Alternator shaft rotational speed
Radial load

500
0

0

0.05

0.1

0.15

Time

0.2

0.25

1400

Photos of authors
Radial load N

Rotational speed min-1

100

1300
1200
0.3

sec

Fig. 9 Results of Idling test

-127-

Koji SATO

Isao MIKURIYA

Automotive Engineering Dept.
Automotive Sales Headquarters

Automotive Engineering Dept.
Automotive Sales Headquarters

NTN TECHNICAL REVIEW No. 75（2007）

Presentation of award-winning articles
Society of Tribologists & Lubrication Engineers Captain Alfred E. Hunt Award

Crack Propagation of Rolling Contact Fatigue
in Ball Bearing Steel Due to Tensile Strain
Noriyuki TSUSHIMA**
Crack propagations or failure modes in rolling element bearings, which had been difficult to explain
via conventional crack propagation mechanisms such as the orthogonal shear stress mechanism,
were discussed from the viewpoint of a tensile strain mechanism. Contact stresses are compressivein three axes, whose values differ from each other, then strain can be tensile in one of these
directions, acting at a right angle to the direction of the maximum compressive stress. A crack is
considered to propagate by this tensile strain. When contact stress is small, a crack produced bysome cause can propagate by this elastic tensile strain. When contact stress is large, residual
tensile strain is produced by plastic deformation, which can also influence the crack propagation.
Several failure modes of rolling element bearings, which had been difficult to explain, were explainedby tensile strain.

principal stresses are all compressive, but a stresssubtracted by hydrostatic pressure can be tensile andthis stress is called relative tensile stress by Schlicht.
Voskamp and Hollox 2) and Voskamp 3) discussed theinfluence of tensile residual stress that has beencreated by plastic deformation under large contact
stress on crack initiation and propagation. Recently,
Harris and McCool 4) considered von Mises equivalentstress and Zaretsky et al.5) considered von Mises
equivalent stress or maximum shear stress to beappropriate because the orthogonal shear stresscannot be used to evaluate the influence of residual
stress on rolling contact fatigue.
On the other hand, in 1967, Lyman 6) discussed therole of tensile strain as the rolling contact fatigue
mechanism, and in 1996, Muro and Yamamoto 7), 8)
tried to explain the mechanisms of rolling contact
fatigue, such as flaking, pitting, and spalling, by tensile
strain as a unified theory for crack propagation under
three principal compressive stresses.
Succeeding Muro's study, the present article
explains the crack propagation of rolling contact
fatigue by tensile strain, using stress and strain
calculation programs developed by Muro 9). As
examples for the clarifying explanation , rolling
element bearing failures, whose involvement in thecrack propagation mechanism has been difficult to
explain, are mentioned.

1. Introduction
There are several theories for crack initiation andpropagation mechanism in rolling contact fatigue. In
the 1940s, Lundberg-Palmgren considered orthogonal
shear stress to be the crack controlling stress because
he observed arrested cracks lying parallel to thesurface on a section of bearings after rolling contact
fatigue. After that, the most commonly used theory
was that the crack initiates and propagates byorthogonal shear stress.
However, as cited in the examples mentioned in thepresent article, there are failures that cannot be
explained by the orthogonal shear stress: crack
initiation and propagation under the existence ofresidual stress, through rolling element fracture or
through inner ring fracture under fitting stress. These
failures are closely related to the amount of residual
stress or fitting stress, but the amount of orthogonal
stress is not influenced by the existence of residual
stress or fitting stress. Therefore, they cannot be
explained by orthogonal stress.
After 1980, various theories emerged. Zwirlein and
Schlicht 1) proposed a concept of relative tensile stress--that is, the stress in a certain direction subtracted byhydrostatic pressure-- discussing its influence on crack
initiation and propagation under the three principal
compressive stress states. In rolling contact, the three*Former Bearing Engineering R&D Center-128-

Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strainstress by multiplying the elastic deviatoric stress by afactor of√3 /σM. This modification for residual stresscalculation was made to stabilize the calculation,
which otherwise had scattered between a given
loading cycle and the next loading cycle, making
saturation difficult (Muro 9)).
The rolling direction is defined as the x-axis
direction, the axial direction is defined as the y-axis
direction, and the depth direction is defined as the zaxis direction, and for example the xy plane is theplane made by x and y axes. The half-radius of the
major axis of contact ellipse is a and the half-radius of
the minor axis is b. Load movement is from x = -2b to+2b .
Stress values obtained are normalized by dividingthem by the yield shear stress k . Strains obtained are
multiplied by Young's modulus to express it in stressunits for the convenience of comparison with thematerial strength, because stress values for thematerial strength are usually available in the literatureand strain values are not. Because the yield shear
stress k of hardened ball bearing steel is about 1 GPa(derived from the fact that one-dimensional yield shear
stress k is one-dimensional tensile yield stress 1.7
GPa divided by √3), then the expressed values in thefollowing tables can be considered values in GPa.

2. Calculation Program
Programs used for stress and strain calculations arethose developed by Muro. The source codes havealready been published (Muro 9)), and the summary of
the code is described in the Appendix. These
programs were changed for their language (Basic to
Fortran) but were not revised in content. The program
calculates the stress at a given point by using theBoussinesq equation for the concentrated force andsums the stresses due to every distributed force
composing Hertzian contact stress. A strain
calculation is done, using the stress obtained as
mentioned earlier, in accordance with the method of
Merwin and Johnson 10). Their method is as follows.
The stress at a given point of calculation increases as
the loading point moves and, when the stresssurpasses the elastic limit of the material, thedeviatoric stress changes according to the PrandtlReuss strain incremental relation. The difference
between elastic and plastic deviatoric stresses is theresulting residual stress. After unloading, the residual
stress in the depth direction is eliminated bytransforming it to residual stresses in the rolling andaxial directions, keeping the strains in the rolling andaxial directions as they are, because residual stress inthe depth direction is assumed to be zero. However, in
Muro's program, the residual stress in the depth
direction is not transformed into residual stresses in
the rolling and axial directions for the next loading,while making the residual stress in depth direction
zero. This procedure was said to stabilize thecalculation. Furthermore, Muro's program assumes
that plastic deformation occurs when von Mises
equivalent stress σM divided by the yield shear stressk exceeds √3 and simply calculates plastic deviatoric

3. Examples of Crack Propagation by
Tensile Strain
In this section, four rolling contact fatigue failures
were cited for the explanation of crack propagation bytensile strain. They were thus far difficult to explain byordinary mechanisms such as orthogonal shear
stress. Specimen types, materials, dimensions, andlubricants are shown in Table 1.

Table 1 Test or running conditions for the examples of crack propagation analysls by tensile strain in this article
ExampleSpecimen,
dimensions (Material)

Matching specimen,
dimensions (Material)

a /b

Pmax
(GPa)

Slip

LublicantRelated figure(Reference)A

Ring (SUJ2)OD : 50mm
AC : 50R

Ring (SUJ2)OD : 53mm
AC : 50R

1.6

3.0
4.6

No

Turbin oil
pad supply

Fig. 1, Fig. 2(Muro et al.11) )B

Ring (SUJ2)OD : 50mm
AC : 25.0R

Ring (SUJ2)OD : 53mm
AC : 26.5R

1.0

5.0

No

Turbin oil
pad supply

Fig. 3, Fig. 4(Muro et al.11) )C

Ball bearing 6303
Outer ring（SUJ2）

9/32 in. Ball（SUJ2）

5.5

2.9

No

GreaseFig. 6(Tamada and Tanaka et al.12) )D

Spherical roller bearing 23160
Inner ring（SUJ3）

Spherical roller（SUJ3） 54.0

2.0

No

GreaseFig. 7(Tsushima et al.13) )OD: Outer diameter, AC: Radius of curvature in axial direction, SUJ2: AISI 52100 Equivalent in JIS-129-

NTN TECHNICAL REVIEW No. 75（2007）

RS1 = 0.4k (that is, 0.4 GPa) in the rolling direction
and RS2 = 0.2k (= 0.2 GPa) in the axial direction. The
values of RS1 and RS2 were estimated from X-ray
measurement in Figs. 2(a) and 2(b). In this test, theellipticity of the contact surface is a/b=1.6. Therefore,
calculation was made with a/b=1.6. The depth was
normalized by b, the half-radius of the minor axis of
the contact ellipse.

Fracture and Flaking of the Ring Specimen
The failure mode of the ring specimen (Muro et al.11))
with 50 mm in diameter after rolling contact fatigue
was flaking irrespective of load values when thespecimen was not given residual tensile stress at thesubsurface beforehand. However, when the specimen
was given tensile residual stress before the test, thefailure mode was fracture when the load was small,and flaking when the load was large. Fig. 1 shows thering specimen fractured without flaking.
As the explanations for these different failuremodes, the following was given by X-ray residual
stress measurement shown in Figs. 2(a) and 2(b).
When the load is small (Fig. 2(a)), the creation ofresidual compressive stress is small and the crack can
propagate in the depth direction. When the load islarge (Fig. 2(b)), compressive residual stress iscreated and the crack propagation into the depth
direction is thus inhibited; then flaking occurs.
However, a quantitative explanation for the crack
propagation was difficult at that time. In the following,
an explanation via tensile strain is attempted.
Tables 2 and 3 show calculated stressesσx, σy, σz
in x-, y- and z-axis directions and maximum shear
stress τ45, orthogonal shear stress τo, and von Mises
equivalent stress σM divided by √3 at the center ofcontact track (x = 0, y = 0) for the ring specimens at
the first loading of rolling contact with Pmax = 3.0k (thatis, 3.0 GPa).
In the test, ring specimens were subjected tostandard heat treatment and some of them werepreviously given tensile residual stress after heat
treatment by induction heating of the subsurface.
Table 2 shows the stresses for the standard heat
treatment specimen. Table 3 shows stresses for the
specimen previously given residual tensile stress of

Rolling direction

10mm

Fig. 1 Fracture of ring specimen tested at Pmax=3GPa11)

Fig. 2 Residual stress distributions of ring specimen
measured by X-ray after testing at(a) Pmax= 3 GPa and (b) Pmax = 4.6 GPa (Muro et al.11))

Table 2 Calculated stresses under contact surface atinitial stage of loading for ring specimen not given
residual stress before test
（Calculating condition : Pmax=3.0GPa, RS1= 0GPa,
RS2=0GPa, a/b=1.6）

Table 3 Calculated stresses under contact surface atinitial stage of loading for ring specimen given residual
stress before test
（Calculating condition : Pmax=3.0GPa, RS1= 0.4GPa,
RS2=0.2GPa, a/b=1.6）

z/b

σx

σy

σz

τo

τ45

σM/√3

z/b

σx

σy

σz

τo

τ45

σM/√3

0.01

-2.52

-2.26

-3.18

0.14

0.33

0.48

0.01

-2.12

-2.06

-3.18

0.14

0.53

0.49

0.1

-1.99

-1.88

-2.98

0.46

0.50

0.61

0.1

-1.59

-1.68

-2.98

0.46

0.69

0.62

0.2

-1.54

-1.54

-2.92

0.62

0.69

0.80

0.2

-1.14

-1.34

-2.92

0.62

0.89

0.80

0.3

-1.17

-1.26

-2.82

0.68

0.83

0.93

0.3

-0.77

-1.06

-2.82

0.68

1.03

0.93

0.4

-0.88

-1.02

-2.70

0.69

0.91

1.02

0.4

-0.48

-0.82

-2.70

0.69

1.11

1.10

0.5

-0.65

-0.82

-2.56

0.68

0.95

1.06

0.5

-0.25

-0.62

-2.56

0.68

1.15

1.05

0.6

-0.48

-0.66

-2.41

0.66

0.96

1.07

0.6

-0.08

-0.46

-2.41

0.66

1.16

1.06

0.7

-0.35

-0.53

-2.25

0.63

0.95

1.05

0.7

0.05

-0.33

-2.25

0.63

1.15

1.05

0.8

-0.25

-0.42

-2.10

0.59

0.92

1.02

0.8

0.15

-0.22

-2.10

0.59

1.12

1.02

0.9

-0.18

-0.33

-1.94

0.56

0.88

0.98

0.9

0.22

-0.13

-1.94

0.56

1.08

0.98

1.0

-0.12

-0.26

-1.80

0.52

0.84

0.93

1.0

0.28

-0.06

-1.80

0.52

1.04

0.93

-130-

Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain

Comparison between Tables 2 and 3 shows thatthe orthogonal shear stress τo is the same for thesame depth despite other stresses being different.
Although the orthogonal shear stress is the sameirrespective of the existence of residual stress, thefailure modes are different as mentioned earlier. Then
it is clear that the orthogonal shear stress does notcontrol crack propagation.
Tables 4(a) and 4(b) show the residual stresscalculations when rolling contact is done with Pmax=
3.0k and 4.6k, respectively. Since residual stressvalues would be different with the axial position (thatis, the y coordinate), the calculation was made at y =0.5a to be able to compare the results with the X-ray
measurement in Fig. 2, which will produce theaverage value over the contact track. RRx is theresidual stress in the rolling direction (x direction) and
RRy is the residual stress in axial direction (y direction);
both are normalized by yield shear stress k in thetables. These are values at the loading cycle whenplastic deformation reached or almost reached thesaturation so that the plastic deformation lessened.
Residual stresses in Tables 4(a) and 4(b) would bealmost similar, respectively, to the residual stressvalues in Figs. 2(a) and 2(b) measured by X-ray.
Table 5 shows calculated "residual" tensile strain
E1zxr at each depth and the maximum value of "elastic"
tensile strain E1zx for x = -2b to +2b, at y = 0.5a. (Notethat E1zxr and E1zx are the principal tensile strains on
the zx plane made by z and x axes.) This maximum
elastic tensile strain exists outside of the contact
surface (that is, at x < -1.0b or x > +1.0b) and thevalue (for example, 0.45 ± 30x) in the table meansthat the elastic tensile strain changes its value during
the movement of loading and the maximum strain 0.45
inclines by +30˚ clockwise from the x axis for the

Table 4 Calculated residual stresses at y=0.5a of ring
specimen given residual stress before calculation(a) Pmax=3k,(b) Pmax=4.6k,
RS1=0.2k, RS2=0.2k
RS1=0.4k, RS2=0.2kz/bz, mm RRx/k RRy/kz/bz, mm RRx/k RRy/krange x < 0 and by -30˚ for x > 0. As mentioned
earlier, the value of the strain is the value multiplied byYoung's modulus to transform it to the stress unit.
Table 5(a) shows that, in the case of Pmax = 3.0k,
the residual tensile strain E1zxr becomes extremely
small (that is, about 0.02k or 0.02 GPa) for a depth
larger than z=0.22 mm as the crack propagates into
the depth. Because the critical value of the stressintensity factor range ΔKIth for the crack propagation
by fatigue is assumed to be 5 MPa √m (Muro 9)), thestress for the crack to propagate in this case is 5÷
√π 0.0002 = 0.2 GPa (∵ΔKI =Δσ√πa), much larger
than the 0.02 GPa mentioned earlier. Then the crack
cannot propagate by residual tensile strain. On theother hand, since elastic tensile strain E1zx is large
enough (over 0.30 GPa), the crack can propagate into
the depth direction by this elastic tensile strain. The
direction of elastic tensile strain changes from -30˚ to+30˚ after z = 0.11 mm as the contact point moves;then the crack is thought to propagate
macroscopically in a zigzag with changing directions.
On the contrary, in the case of Pmax = 4.6k in Table
5(b) even around z = 0.25 mm, residual tensile strain
E1zxr is large, about 0.30 GPa, and a crack can
propagate by this strain. In this case, residual tensile
strain directs in the z direction (denoted by 0z in thetable), and then the crack propagates at a right angle
to this; namely, parallel to the surface or xy plane,
finally resulting in flaking. Cracks can also propagate
by elastic tensile strain E1zx because it is about 0.5
GPa, but the direction changes variably as the loading
point moves; then its contribution would be small.
Therefore, it is considered that the crack propagatesparallel to the surface mainly by residual tensile strain
E1zxr.

Table 5 Calculated residual E1zxr, and elastic principal strains E1zx at
y=0.5a of ring specimen given residual stress before calculation(Note: E1zxr, and E1zx are principal tensil strains on zx plane.)(a) Pmax=3k,(b) Pmax=4.6k,
RS1=0.4k, RS2=0.2k
RS1=0.4k, RS2=0.2k
z/b z, mm

E1zxr/kE1zx/kz/b z, mm

E1zxr/kE1zx/k

0.01 0.005

0.40

0.20

0.01 0.008

0.40

0.20

0.01 0.005 0.35 0x 0.53 0x

0.01 0.008 0.35 0x 0.63 0x

0.1

0.40

0.20

0.1

0.35

0.15

0.1

0.05

0.35 0x 0.42 0x

0.1

0.08

0.31 0x 0.43 ±45x0.110.35 0x 0.45 ±30x

0.2

0.17

0.12 0z

0.08

±

0.05

0.2

0.11

0.40

0.20

0.2

0.17

-0.23 -0.25

0.2

0.3

0.16

0.30

0.14

0.3

0.25

-0.59 -0.58

0.3

0.16

0.26 0x 0.44 ±31x

0.3

0.25

0.29 0z

0.50 27z

0.4

0.22

0.02 -0.02

0.4

0.33

-0.59 -0.59

0.4

0.22

0.02 0x 0.33 ±36x

0.4

0.33

0.30 0z

0.48 25z

0.5

0.27

-0.08 -0.06

0.5

0.42

-0.65 -0.63

0.5

0.27

0.04 0z

0.29 ±36x

0.5

0.42

0.32 0z

0.46 24z

0.6

0.32

-0.05 -0.03

0.6

0.50

-0.63 -0.63

0.6

0.32

0.02 0z

0.30 ±29x

0.6

0.50

0.31 0z0.42 27z

0.7

0.38

0.07

0.7

0.58

-0.91 -0.68

0.7

0.38

0.06 0x 0.38 ±18x

0.7

0.58

0.40 0z0.45 12z0.43

0.19 0x 0.52 0x

0.8

0.66

0.26 0z

0.33 22z

±
±
±
±
±
±

0.03

0.45 36z

0.8

0.43

0.22

0.11

0.8

0.66

-0.60 -0.46

0.8

0.9

0.49

0.35

0.18

0.9

0.75

-0.34 -0.27

0.9

0.49

0.31 0x 0.64 0x

0.9

0.75

0.16 0z

0.31 ±42x

1.0

0.54

0.40

0.20

1.0

0.83

-0.35 -0.25

1.0

0.54

0.35 0x 0.68 0x

1.0

0.83

0.14 0z

0.29 ±36x-131-

NTN TECHNICAL REVIEW No. 75（2007）

the residual compressive stress at the depth of z =0.15b was calculated at first for each y value, then thiskind of residual tensile stress at the depth of z = 0.05b
was assumed by giving the opposite sign to theresidual compressive stress value at the depth of z =0.15b.
Table 6 shows the elastic principal maximum andminimum strains E1xy, E2xy and residual principal
maximum and minimum strains E1xyr, E2xyr at z= 0.05b
on the xy plane parallel to the surface after plastic
deformation reaches saturation. Some representativevalues in Table 6 are plotted in Fig. 5. The maximum
elastic tensile strain exists at y= 0.7a (= 2a/3) and
x=±1.2b. Then we can understand that the crack
initiates at the position y= 2a/3. Namely, in the case of
circular contact, the crack initiation site is not at thecenter of the contact track but at y= 2a/3. On the other
hand, this elastic tensile strain inclines by +36˚ or -36˚
to the x axis at y= 0.7a and x= ±1.2b. Namely, thecrack can propagate at a right angle to this strain
direction in both directions,±54˚to the x axis. In Fig.
3, the crack propagates nearly with this angle but in
one direction. The crack propagation direction isconsidered to be influenced by other factors, forexample, the tangential force. It has been confirmedthat the crack propagation directions were opposite
each other between the upper and lower specimens
when there was a small slip (Muro et al.11)).

Crack Propagation for Circular Contact
In the rolling contact test of the ring specimen(Muro, et al.11)) with 50 mm diameter and with certain
curvature in the axial direction, when the contact
surface is circular (that is, a/b = 1), the crack initiates
not at the center of the contact track but at 2a/3 awayfrom the center of the contact track. Fig. 3 shows thissituation: A crack initiated 2a/3 away from the centerof the contact track and then propagated at about 45˚
to the rolling direction. The crack propagating toward
the center of the contact track turned its direction to
the rolling direction after reaching the center. The
crack also propagated parallel to the surface andresulted in flaking. In the following, the crack initiation
at 2a/3 away from the center of the contact track is
discussed from the viewpoint of tensile strain.
Fig. 4 shows the result of X-ray measurement on
the contact track after the test with Pmax = 5.0 GPa.
Residual tensile stress is created in the surface layer,balancing the created residual compressive stressinside. The calculation program used in the present
article cannot calculate this kind of tensile residual
stress, which is created in a plastically nondeformed
shallower zone for the balance of createdcompressive residual stress at a plastically deformed
deeper zone, because calculation at a given depth isindependent of other upper or lower depths (that is, itdoes not influence the residual stress of other depths).
The residual tensile stress of this kind was given
beforehand in the calculation. Since this kind ofresidual tensile stress of the surface layer isconsidered to be different along the axial position y,depending on the residual compressive stress createdin the neighbor inside (or, in other words, since thelarger the residual compressive stress in deeper
depth, the larger this kind of residual tensile stress),

Abnormal Flaking of Alternator Bearing
An alternator bearing outer ring flakes as shown in
Fig. 6. It is a small flaking in appearance but asectional view reveals many long cracks running in thedepth direction (Tamada and Tanaka12)). Compared to

Track edge

Track center

Moving direction of loading point

Fig. 3 Crack initiation and propagation on the surface incircular contact; arrow shows origin of cracking(Muro et al.11)).

Fig. 4 Residual stress distribution of ring specimen
measured by X-ray (Muro et al.11))-132-

Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain

Table 6 Residual and elastic principal strains E1xyr, E1xy and their directions on xy-plane at
z=0.05b for circular contact, E1 and E2 mean maximum and minimum principal strains,respectively (Calculating condition: Pmax=5.0k, a/b=1)0ayy―direction
a

0.3a

0.7a

1.0a

x

E1xy/kE2xy/kE1xy/kE2xy/kE1xy/kE2xy/kE1xy/kE2xy/k

-2.0b0.14 0x-0.18

0.31 7x-0.05

0.43 29x

0.19

0.14 27x

-0.15

-1.8b

0.16 0x-0.22

0.34 8x-0.09

0.46 30x

0.16

0.16 29x

-0.18

-1.6b

0.20 0x-0.28

0.37 10x-0.14

0.49 31x

0.11

0.18 32x

-0.21

-1.4b

0.23 0x-0.36

0.40 11x-0.22

0.52 33x

0.06

0.21 36x

-0.25

-1.2b

0.21 0x-0.47

0.39 13x-0.32

0.56 36x-0.02

0.23 40x

-0.30

-1.0b

-0.20 0x

-0.64

0.06 16x-0.46

0.55 40x-0.13

0.26 -45y

-0.36

-0.8b

-0.52 0x

-0.96

-0.19 20x-0.76

0.31 -42y-0.26

0.26 -39y

-0.43

-0.6b

-0.96 0x

-1.17

-0.66 25x

-0.98

0.20 -35y-0.49

0.21 -31y

-0.50

-0.4b

-1.23 0x-1.29

-0.94 34x

-1.01

-0.11 -24y

-0.68

0.04 -22y

-0.57

-0.2b

-1.36 0y-1.37

-1.10 -38y

-1.17

-0.29 -12y

-0.79

-0.17 -11y

-0.63

0b

-1.38 0y-1.42

-1.15 8y-1.19

-0.35 0y

-0.83

-0.18 0y

-0.66
-0.63

0.2b

-1.36 0y

-1.37

-1.09 40y-1.18

-0.29 12y-0.79

-0.17 11y0.4b-1.23 0x-1.29

-0.93 -35x

-1.11

-0.11 24y

-0.68

0.04 22y

-0.57

0.6b

-0.96 0x-1.17

-0.65 -26x

-0.98

0.20 35y-0.49

0.21 31y

-0.50

0.8b

-0.52 0x

-0.96

-0.18 -21x

-0.77

0.31 42y-0.26

0.26 39y

-0.43

1.0b

-0.20 0x

-0.64

0.06 -17x-0.47

0.55 -40x-0.13

0.26 45y

-0.36

1.2b

0.21 0x-0.47

0.40 -14x-0.32

0.56 -36x-0.02

0.23 -40x

-0.30

1.4b

0.23 0x-0.36

0.40 -12x-0.22

0.52 -33x

0.06

0.21 -36x

-0.25

1.6b

0.20 0x-0.28

0.37 -11x-0.15

0.49 -31x

0.11

0.18 -32x

-0.21

1.8b

0.16 0x-0.22

0.34 -10x-0.09

0.46 -30x

0.16

0.16 -29x

-0.18

2.0b

0.14 0x-0.18

0.31 -9x-0.05

0.43 -29x

0.19

0.14 -27x-0.15

E1xyr/kE2xyr/kE1xyr/kE2xyr/kE1xyr/kE2xyr/kE1xyr/kE2xyr/k0.01 0y-0.03

0.15 -19y

0.13

0.39

0.26

0

0y

45y

0

Residual principal

Elastic principal strain E1xy/kstrain E1xyr/kb
x―direction

Fig. 5 Residual and elastic principal strains and their directions on xy plane at z = 0.05b shown in Table 6. (Large dotted circle
is circular contact surface. Dense dark lines mean elastic principal maximum strain E1xy/k or residual principal maximum strain
E1xyr/k, which are tensile, and faint blue lines mean elastic principal maximum strain E1xy/k, which are compressive.)

Fig. 6 Cracking on circumferential section of analternator bearing outer ring (Tamada and Tanaka12))-133-

NTN TECHNICAL REVIEW No. 75（2007）

Table 7 Calculated principal strains and their directions
for alternator outer ring（caluculation condition : Pmax=2.9k, y=0, a/b=5.5）
x

0b

-0.2b
+0.2b

-0.6b
+0.6b

x

-1.0b
+1.0b

0.1b

-1.04 ± 0x

-1.00 ± 5x-0.7 ±17x

0.26

0.2b

-0.65 ± 0x-0.61 ± 6x-0.3 ±20x

0.32

z

0b

-0.2b
+0.2b

-0.6b
+0.6b

-1.0b
+1.0b

-1.2b
+1.2b

-1.6b
+1.6b

±

z

Table 8 Calculated shear stress τo acting at right angle
to the surface for the alternator outer ring（caluculation condition : Pmax=2.9k, y=0, a/b=5.5）

37z

0.1b

±0

±0.06

±0.21

±0.42

±0.08

±0.01

0.2b

±0

±0.11

±0.38

±0.55

±0.24

±0.05

±

40z

0.3b

-0.34 ± 0x-0.30 ± 6x-0.0 ±21x

0.36

±

43z

0.3b

±0

±0.15

±0.50

±0.62

±0.37

±0.11

0.4b

-0.09 ± 0x-0.07 ± 7x0.16 ±22x0.38 ±44x

0.4b

±0

±0.19

±0.56

±0.66

±0.46

±0.17

0.5b

0.09 ± 0x0.11 ± 7x0.28 ±22x0.40 ±42x

0.5b

±0

±0.21

±0.60

±0.68

±0.52

±0.23

0.6b

0.22 ± 0x0.24 ± 7x0.35 ±22x0.41 ±40x

0.6b

±0

±0.22

±0.60

±0.68

±0.55

±0.28

0.7b

0.31 ± 0x0.33 ± 7x0.40 ±22x0.42 ±38x

0.7b

±0

±0.22

±0.59

±0.67

±0.57

±0.32

0.8b

0.38 ± 0x0.39 ± 7x0.43 ±21x0.42 ±36x

0.8b

±0

±0.22

±0.57

±0.66

±0.58

±0.36

0.9b

0.42 ± 0x0.42 ± 7x0.44 ±20x0.42 ±34x

0.9b

±0

±0.21

±0.55

±0.64

±0.58

±0.39

1.0b

0.45 ± 0x0.45 ± 6x0.45 ±20x0.42 ±33x

1.0b

±0

±0.20

±0.52

±0.61

±0.57

±0.41

1.2b

0.46 ± 0x0.46 ± 6x0.45 ±18x0.41 ±30x

1.2b

±0

±0.18

±0.46

±0.56

±0.54

±0.42

1.4b

0.46 ± 0x0.45 ± 6x0.44 ±17x0.39 ±28x

1.4b

±0

±0.15

±0.40

±0.51

±0.50

±0.42

1.6b

0.44 ± 0x0.43 ± 5x0.41 ±16x0.37 ±25x

1.6b

±0

±0.13

±0.35

±0.45

±0.46

±0.41

1.8b

0.41 ± 0x0.41 ± 5x0.39 ±14x0.35 ±24x

1.8b

±0

±0.11

±0.30

±0.40

±0.42

±0.39

2.0b

0.38 ± 0x0.38 ± 5x0.37 ±13x0.33 ±22x

2.0b

±0

±0.10

±0.26

±0.36

±0.38

±0.37

2.5b

0.32 ± 0x0.32 ± 4x0.31 ±11x0.28 ±19x

2.5b

±0

±0.07

±0.19

±0.27

±0.29

±0.31

3.0b

0.27 ± 0x0.26 ± 3x0.26 ±10x0.24 ±16x

3.0b

±0

±0.05

±0.14

±0.20

±0.22

±0.25

4.0b

0.19 ± 0x0.19 ± 3x0.18 ± 8x0.18 ±13x

4.0b

±0

±0.03

±0.08

±0.12

±0.14

±0.16

5.0b

0.14 ± 0x0.14 ± 2x0.13 ± 6x0.13 ±10x

5.0b

±0

±0.02

±0.05

±0.07

±0.09

±0.11

Table 9 Calculated principal strains and their directions
for spherical roller bearing inner ring（caluculation condition : Pmax=2.0k, RS1=0.1k, y=0, a/b=54）
z

x

0b

-0.2b
+0.2b

-0.6b
+0.6b

Table 10 Calculated shear stress τo acting at right angle
to the surface for spherical roller bearing inner ring（caluculation condition : Pmax=2.0k, RS1=0.1k, y=0, a/b=54）
x

-1.0b
+1.0b

z

0b

-0.2b
+0.2b

-0.6b
+0.6b

-1.0b
+1.0b

-1.2b
+1.2b

-1.6b
+1.6b

0.1b

-0.73 ± 0x-0.71 ± 5x-0.5 ±17x

0.16

±

37z

0.1b

±0

± 0.04± 0.14± 0.29± 0.06± 0.010.2b-0.46 ± 0x-0.43 ± 6x-0.2 ±20x

0.21

±

40z

0.2b

±0

± 0.08± 0.26± 0.38± 0.17± 0.040.4b-0.06 ± 0x-0.04 ± 7x0.12 ±22x0.26 ±44x

0.4b

±0

± 0.13± 0.39± 0.46± 0.32± 0.120.6b

0.17 ± 0x0.18 ± 7x0.26 ±22x0.30 ±39x

0.6b

±0

± 0.15± 0.42± 0.48± 0.39± 0.200.8b0.29 ± 0x0.30 ± 7x0.33 ±21x0.32 ±35x

0.8b

±0

± 0.15± 0.40± 0.46± 0.41± 0.261.0b0.35 ± 0x0.35 ± 6x0.35 ±19x0.33 ±32x

1.0b

±0

± 0.14± 0.36± 0.43± 0.41± 0.291.5b0.38 ± 0x0.38 ± 5x0.36 ±16x0.33 ±25x

1.5b

±0

± 0.10± 0.27± 0.35± 0.35± 0.312.0b0.36 ± 0x0.36 ± 4x0.34 ±13x0.31 ±21x

2.0b

±0

± 0.07± 0.19± 0.27± 0.28± 0.282.5b0.33 ± 0x0.33 ± 4x0.32 ±11x0.30 ±17x

2.5b

±0

± 0.05± 0.14± 0.20± 0.22± 0.243.0b0.30 ± 0x0.30 ± 3x0.29 ± 9x0.28 ±14x

3.0b

±0

± 0.04± 0.11± 0.16± 0.18± 0.203.5b0.28 ± 0x0.28 ± 3x0.27 ± 8x0.26 ±12x

3.5b

±0

± 0.03± 0.08± 0.13± 0.14± 0.174.0b0.26 ± 0x0.26 ± 2x0.26 ± 7x0.25 ±11x

4.0b

±0

± 0.02± 0.07± 0.10± 0.12± 0.145.0b0.23 ± 0x0.23 ± 2x0.23 ± 5x0.23 ± 9x

5.0b

±0

± 0.01± 0.04± 0.07± 0.08± 0.106.0b0.21 ± 0x0.21 ± 1x0.21 ± 4x0.21 ± 7x

6.0b

±0

± 0.01± 0.03± 0.05± 0.06± 0.078.0b

0.19 ± 0x0.19 ± 1x0.19 ± 3x0.19 ± 5x

8.0b

±0

± 0.01± 0.02± 0.03± 0.03± 0.0410.0b

0.17 ± 0x0.17 ± 1x0.17 ± 2x0.17 ± 4x

10.0b

±0

± 0.00± 0.01± 0.02± 0.02± 0.0312.0b0.16 ± 0x0.16 ± 1x0.16 ± 2x0.16 ± 3x

12.0b

±0

± 0.00± 0.01± 0.01± 0.02± 0.0214.0b0.15 ± 0x0.15 ± 0x0.15 ± 1x0.15 ± 2x

14.0b

±0

± 0.00± 0.01± 0.01± 0.01± 0.0216.0b0.14 ± 0x0.14 ± 0x0.14 ± 1x0.14 ± 2x

16.0b

±0

± 0.00± 0.00± 0.01± 0.01± 0.0120.0b0.13 ± 0x0.13 ± 0x0.13 ± 1x0.13 ± 1x

20.0b

±0

± 0.00± 0.00± 0.00± 0.01± 0.01-134-

Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain

ordinary flaking, it is characterized by cracking into thedepth direction, sometimes deeper than 0.6 mm (z=2.5b) or rarely deeper than 0.8 mm (z= 3b), and themechanism has not been clear. This crack propagation
into the depth direction is discussed later from theviewpoint of tensile strain. In this case, the cause of
the failure has been clarified: the hydrogen producedby the decomposition of the lubricant influenced thiskind of failure (Tamada and Tanaka12)).
Ellipticity of the contact surface between the outerring and the ball is a/b= 5.5. A test was conductedwith Pmax= 2.9 GPa. Elastic tensile strains and theirdirection angles calculated for each depth at Pmax=
2.9k are shown in Table 7. In Tables 7 to 10, forexample, a value such as -1.00±5x at x= -0.2b and+0.2b for z= 0.1b in Table 7 shows that at x= -0.2b theprincipal strain is -1.00k with clockwise inclination of+5˚ to the x axis and at x= +0.2b the principal strain is-1.00k with clockwise inclination of -5˚ to the x axis. It
is considered that, below the depth z= 1.0b (= 0.25mm), tensile strain in the x direction, namely parallel tothe surface, is a maximum and the value transformed
to stress unit is 0.45k (= 0.45 GPa). Even at z = 2.5b(= 0.6 mm), there exists tensile strain corresponding to0.32 GPa. Due to material deterioration by hydrogen,it will be possible for the crack to initiate andpropagate at a right angle to the x axis, namely into
the depth direction, by this strain value.
On the other hand, orthogonal shear stress acting at
a right angle to the surface is shown in Table 8. The
orthogonal shear stress around the depth z = 1.0b (=0.25 mm) is about 0.6 GPa at x = -1.0b and + 1.0b but
it decreases rapidly as the depth increases. Below z=
2.5b (= 0.6 mm), it is smaller than 0.3 GPa. In thecase of high-hardness steels, for the crack topropagate it is said that the shear stress value must
be three times larger than the tensile stress value.
Then it would be difficult for the crack to initiate andpropagate by this shear stress value. It will bereasonable to consider that the crack propagated bytensile strain.

Fracture of Spherical Roller Bearing Inner Ring
Some type of spherical roller bearing inner ring isused under some value of fitting stress. When thefitting stress is larger than 0.1 GPa, fracture of theinner ring rarely happens during service. It isconsidered that the fracture is caused by the repetitionof small tensions due to fitting stress and large
compression due to contact stress (Tsushima, et al.13)).
Fig. 7 shows the fracture surface. The crack initiatednear the surface of z= 0.05b (= 0.04 mm) andpropagated macroscopically in zigzag formation into
the depth until z= 4b (= 3 mm); then it propagated
gradually, creating a smooth fracture surface andrapidly from z= 20b (= 16 mm). Macroscopically zigzag
crack propagations to a depth of 3 mm and smoothcrack propagation from 3 to 16 mm have not yet beenclarified. In the following, this crack propagation
mechanism is discussed from the tensile strain
viewpoint.
The ellipticity of the contact surface between the
inner ring and roller is a/b= 54. We assume Pmax= 2.0k(that is, 2.0 GPa) as the Hertzian contact stress duringservice. Calculation of tensile strain and its direction isshown in Table 9. Here, stress of 0.1k (= 0.1 GPa)
was given as the fitting stress RS1 in thecircumferential direction (x direction). At certain
depths, for example, at z= 0.6b (= 0.45 mm) or
deeper, as the loading point moves (x increases fromx= -1b to +1b), tensile strain changes its direction
continuously from over +10ｰ to below -10˚ against thex axis up to the depth z= 4b (= 3 mm). Then, in thisdepth range, the crack is considered to propagate,
sometimes changing its direction, which creates themacroscopic zigzag appearance of the fracturesurface. Below the depth z= 4b (= 3 mm), tensile strain
is within 10˚ against the x axis and the crack is
considered to propagate into the depth, creating asmooth fracture surface. Shear stress acting parallelto the depth direction, shown in Table 10, became
significantly small below z= 4b (= 3 mm) and would not
be able to propagate the crack.
According to fracture mechanics, a crack can
propagate by fatigue when the stress intensity factor
range ΔKI at the crack tip exceeds the critical value ofthe material ΔKIth. For a hardened ball bearing steel,
DKIth is 5 MPa √m (Tsushima et al.13)). The stressintensity factor range ΔKI is roughly Δσ√πa (a iscrack depth); here Δσ is stress range. In this case,stress changes from a value in Table 9 during loading
to the value of fitting stress 0.1 GPa during unloading.
Then Δσis the value in Table 9 subtracted by 0.1
GPa.
Now we examine whether the crack can propagate
from 3 mm (z= 4b) to 16 mm (z= 20b). From Table 9,

Fig. 7 Fracture of a spherical roller bearing inner ring
(Tsushima et al.13）). The arrow shows the origin of the cracking-135-

NTN TECHNICAL REVIEW No. 75（2007）

other than rolling contact fatigue, such experiments
are expected at present.

we know that tensile stress at z= 4b converted fromstrain is about 0.26 GPa. Therefore, the stress range
Δσat z= 4b is 0.16 GPa = 160 MPa. Then at z= 4b or
0.003 m depth, ΔKI is 160 √π0.003 = 15.5 MPa √m,which is larger than the critical value of 5 MPa √m.
Then the crack can propagate by fatigue. Similarly, at
z= 20 or 0.016 m, the stress range is 0.03 GPa = 30
MPa. Then ΔKI =√π0.016 = 6.7 MPa √m, which islarger than the critical value of 5 MPa √m. Then acrack can propagate by fatigue at least until thisdepth. At 0.016 m depth, however, stress intensity
factor KI due to fitting stress σ√πa=100 √π×0.016 =22.4 MPa √m, reaching the fracture toughness value
KIC of this material (estimated as 21 MPa √m by
Tsushima et al.13)), then rapid fracture occurred.

5. Conclusion
So far, as the mechanism of rolling contact fatigue,
orthogonal shear stress acting parallel to the surface
is considered to be the controlling stress for crack
initiation and propagation. However, in failure
analyses of rolling element bearings, there are somecases that cannot be explained by the orthogonal
shear stress. In particular, they are phenomena inwhich a crack propagates deeply into the depth
direction, such as through fracture of an inner ringunder fitting stress. Contact stress is compressive inthree axes, but the values are different; then strain
can be tensile in the direction at a right angle to the
maximum-compression stress direction. We consider
that the crack propagates by this tensile strain. Whencontact stress is small, a crack, produced by somecause, can propagate by this elastic tensile strain.
When contact stress is large, residual tensile strain is
produced by plastic deformation and this residual
tensile strain can also influence the crack propagation.

4. Discussion
So far, the authors have reported the influence ofheat treatment stress (Maeda and Tsushima14)), fitting
stress (Maeda and Tsushima14)), and residual stressdue to mechanical treatment on rolling contact fatigue
life (Maeda et al.15)). It has been proved that, if small
tensile stress such as 0.2 GPa exists, a crack can
initiate during rolling contact fatigue due to this tensilestress and compressive stress due to contact stress-that is, by the repetition of small tension and large
compression (Muro et al.11)). However, as mentioned
in the present article, there are some phenomena
where cracks are propagated without such tensilestress. In these cases, however, by considering thatthe tensile strain is equivalent to tensile stress, theexplanation of crack initiation and propagation ispossible. The contact stress is compressive in threeaxis directions, but their values are not the same; thentensile strain can be produced in any of the three
directions. This tensile strain can have enough value
for crack propagation. Noticing this tensile strain, we
can explain these phenomena that have been difficult
to explain.
As fracture phenomena of hard and brittle materials
by compression, there are some papers that reported
crack initiation parallel to the direction of compression.
For example, a cylinder specimen made from ballbearing steel rarely fractures vertically by compression(Muro16)), or a cylinder specimen made from marble
fractures circumferentially by lateral compression(Sato et al.17)). Vertical fracture of a knock pin of a cold
forging tool by fatigue was also reported (Hamatani 18)).
In these papers, the mechanism of crack propagation
was not mentioned, but it may be considered that thecracking is due to tensile strain, as pointed out by
Muro9). Since papers are few that reported fatigue
crack propagation by tensile strain using a method

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Fatigue Mechanism-Accelerated Testing Versus Field
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2) Voskamp, A. P. and Hollox, G. E. (1988), "Failsafe
Rating of Ball Bearing Components," ASTM STP 987,
p 102.
3) Voskamp, A. P. (1997), "Microstructural Changes
during Rolling Contact Fatigue," Ph. D. Thesis, Delft
University of Technology, Delft, p 184.
4) Harris, T. A. and McCool, J. I. (1995), "On the
Accuracy of Rolling Bearing Fatigue Life Prediction,"
ASME 95-TRIB-31.
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JAST Tribology Conference, Kitakyushu, p 25.
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Application to Bearing Failure Analysis," J. Jpn. Soc.
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NTN TECHNICAL REVIEW No. 75（2007）

Appendix: Calculation Program of Residual Stress by MUROWhen the next stress P2 is applied, those residual
stresses RSx1, ..., RSxy1, which have been created by
Eqs. [A1]--[A6] are added to the stresses of P2, andanother residual stress is created as in Eqs. [A7]-[A12] ifσM2 > √3.:
RSx2=sx2'−sx2=sx2×（√3/σM2）−sx2 ………………[A7]RSy2=sy2'−sy2=sy2×（√3/σM2）−sy2 ………………[A8]RSz2=sz2'−sz2=sz2×（√3/σM2）−sz2 ………………[A9]RSxz2=τxz2'−τxz2=τxz2×（√3/σM2）−τxz2 ………[A10]RSyz2=τyz2'−τyz2=τyz2×（√3/σM2）−τyz2 ………[A11]
RSxy2=τxy2'−τxy2=τxy2×（√3/σM2）−τxy2 ……[A12]After the final stress P41 is applied, created residual
stresses are RSx41, RSy41, RSz41, RSxz41, RSyz41, andRSxy41.
Until this procedure, the method is almost the sameas Merwin and Johnson's10) except for the plastic
deviatoric stress equation. Johnson used PrandtlReuss incremental theory for obtaining plastic
deviatoric stress, which requires solution of differential
equations. Solving differential equations led toconsiderable scatter of calculated values. Therefore,
Muro used the simple method of multiplying the elastic
deviatoric stress by a ratio of √3/σM by the elastic
deviatoric stress to obtain the plastic deviatoric stress.
According to Johnson, after unloading, the residual
stress in the depth direction RSz41 is converted to RSx41
and RSy41 in the x and y directions by using thefollowing equations, because no residual stress exists
in the depth direction, keeping the strains in x and yaxes the same:（RSx41）=RSx41−RSz41×ν/（1−ν）…………[A13]（RSy41）=RSy41−RSz41×ν/（1−ν）…………[A14]（RSz41）=0 ………………………………………[A15]where RSx41, RSy41, and RSz41 are new residual
stresses after conversion, andνis Poisson's ratio.
Residual strains are calculated from residual stresses.
Different from Johnson, Muro's program does not
use this conversion of Eqs. [A13] and [A14] for the
next loading, while using Eq. [A15]. In that case, it
may deviate a little from theoretical values. When Eqs.
[A13] and [A14] were used, calculated residual
stresses were scattered significantly between values
at a given loading cycle and the next loading cycle.
Because RSx41 and RSy41 are negative and RSz41 is
positive in Eqs. [A13] and [A14], not using Eqs. [A13]and [A14] means that smaller residual compressive
stresses are given for the next loading. This stabilizes
the residual stress calculation.

Residual stress and strain calculations in this article
were performed using Muro's program9). Here Muro's
calculation is briefly described.
We assume that the stress distribution along therolling direction from x= -2b to +2b i. e., at 41 points,
x= x1, x2, ..., x41 apart from each other with 0.1b
distance for a given depth isx1（σx1，σy1，σz1，τxz1，τyz1，τxy1），
x1（σx2，σy2，σz2，τxz2，τyz2，τxy2），…
xn（σxn，σyn，σzn，σxzn，σyzn，σxyn），…
x41（σx41，σy41，σz41，τxz41，τyz41，τxy41）
where σx,σy, andσz are normal stresses in x, y,
and z directions, respectively, andτxz, τyz, andτxy are
shear stresses on xz, yz, and xy planes, respectively.
Then the von Mises equivalent stressσM at a point xn
is
σMn = { （[ σxn−σyn）2+（σyn−σzn）2+（σzn−σxn）2+6（τxzn2+τyzn2+τxyn2）]÷2 }1/2
Then we assume that during rolling contact, at agiven point, contact stresses P1, P2,..., P41 aresuccessively applied; here P1, P2,..., P41 correspond to
the stresses of x1, x2, ..., x41, respectively.
Stressesσxn,σyn andσzn at a given point are divided
into respective elastic deviatoric stresses sxn, syn, szn,
and hydrostatic pressure sn. Thus, σxn=sxn+sn, σyn=
syn+sn, σzn=szn+sn, where sn=（σxn+σyn+σzn）/3.
Therefore,sxn=σxn−sn, syn=σyn−sn, szn=σzn−sn. Then
we assume that, if σMn is greater than √3, elastic
deviatoric stresses sxn, syn, szn and shear stressesτxzn,
τyzn, τxyn change by plastic deformation, dependingon the value ofσMn , to sxn×√3/σMn, syn×√3/σMn, ...,τ
3/σMn. Residual stress is the difference between
xyn×√plastic deviatoric stress and elastic deviatoric stress.
Then, at a given point, when applied with a stressP1, if σM1 is greater than √3, residual stress is createdand each residual stressRSx1, ..., RSxy1 is calculated asfollows:
RSx1=sx1'−sx1=sx1×（√3/σM1）−sx1 ………………[A1]
RSy1=sy1'−sy1=sy1×（√3/σM1）−sy1 ………………[A2]RSz1=sz1'−sz1=sz1×（√3/σM1）−sz1 ………………[A3]RSxz1=τxz1'−τxz1=τxz1×（√3/σM1）−τxz1 ………[A4]RSyz1=τyz1'−τyz1=τyz1×（√3/σM1）−τyz1 ………[A5]RSxy1=τxy1'−τxy1=τxy1×（√3/σM1）−τxy1 ………[A6]where sx1, sy2, ...,τxy1 are elastic deviatoric stresses
and sx1', sy1', ...,τxy1' are plastic deviatoric stresses.
RSx1, RSy1, ...,τRSxy1are added to σx1, σy1, ...,τxy1,
respectively, and if this newσM1 is greater than √3,
calculation of Eqs. [A1]--[A6] is repeated until newσM1
becomes smaller than √3.
-138-Crack Propagation of Rolling Contact Fatigue in Ball Bearing Steel Due to Tensile Strain

Author's Comment
It is very important for the simulation of aphenomenon that the calculating value asymptoticallyconverges to one value. Plastic deformation includes
a wide variety of many factors such as yield point,
work hardening, change of contact stress due toplastic deformation, material flow, creation of residual
stress, and loading cycles. Therefore, it is very difficult
to express it in an exact formulation. The author thinksthat Muro's residual stress calculation method used inthis article can produce reasonable values similar to
the actual value as stated in the text (compare Table
4 and Fig. 2).
There may be other methods to calculate residual
stress by rolling contact, such as the finite element
method. However, as just mentioned, it is not easy tosimulate exactly a phenomenon that is affected bymany factors even if we used a precise tool. Muro's
program can be performed on an ordinary personal
computer. In this meaning, the author thinks such a
program is very useful.

Author who gets Hunt Awald
from STLE President, D. McCoy

Photo of author

Noriyuki TSUSHIMA
Formar Bearing Engineering
R&D Center-139-

NTN TECHNICAL REVIEW No. 75（2007）

Presentation of award-winning articles
The 2006 Japan Society of Mechanical Engineers Young Engineers Award

Logarithmic Profiles of Rollers in Roller Bearings
and Optimization of the Profiles*
Hiroki FUJIWARA**
Tatsuo KAWASE**
When a bearing roller is in contact with raceways, excessive pressure peaks occur at the ends of thecontact rectangles. This is called edge loading. Roller and/or raceway profiles are usually crowned toprevent edge loads. Lundberg developed a logarithmic function for a crowned profile. The profile
gives an axially uniform pressure distribution. Johns-Gohar improved the function for the convenienceof manufacturing. However, the Johns-Gohar profile yields edge loading when the roller is tilted. In
addition, the profile allows no straight portion on the roller surface although it is desirable to have a flat
region from the viewpoint of machining. In this study, we modified the Johns-Gohar logarithmic
function to exclude edge loading even when the roller is tilted, allowing a flat region. Threeparameters, K1, K2 and zm, are introduced into the Johns-Gohar function. K1 is coefficient of load, K2 isratio of crowning length to effective contact length, and zm is crown drop at edge of effective contact
length zone. In addition, a mathematical optimization method is used to efficiently determine a set ofthe parameters. An optimization problem is considered to minimize the maximum contact pressure
Pmax , or to maximize the rolling fatigue life L10. A Rosenbrock method is also adopted as the
optimization algorithm. This method requires no evaluation of gradients of the objective function.
Pressure distribution is calculated by making use of a multilevel method. Some examples aredemonstrated to verify the proposed method for both Pmax and L10.

pressure can be made axially uniform. However, thiscurve has an infinite drop at the end of the effective
contact length.
Lundberg gives a finite value as an approximation to
the drop at the end. However, this method has its
disadvantageous in that the shape becomes
discontinuous depending on design conditions. JohnsGohar presented a logarithmic functional formulaobtained by improving Lundberg's method. 2)
Incidentally, the housing and the shaft of a roller
bearing are not exactly parallel to each other, with acertain misalignment present. In addition, themisalignment varies depending on the deflection of
the shaft caused by the loading condition. Under these
conditions, the roller rotates about the x-axis as shownin Fig. 1 and slants against the inner or outer ring; thisis called tilt. In the Johns-Gohar curve, the tilt of aroller causes edge stress to occur.
Reusner states that forming a crowned part in a

1. Introduction
It is known that when a cylindrical surface or conicalsurface comes in contact, concentrated stress occurs
at the end portion of the contact with the resultant
contact pressure becoming excessive. This excessive
contact pressure at the end portions is called edge
stress. In a common roller bearing the roller rollingsurface and/or the raceway surface of the race iscrowned to avoid edge stress. Fig. 1 schematicallyshows a cylindrical roller bearing whose roller iscrowned. The amount of decrease in the radius
generated by crowning is called the drop.
Shapes of crowning include a straight line, a single
circular arc or a combination of multiple circular arcs.
However, Lundberg worked out a shape of crowning
expressed in a logarithmic function (hereinafterreferred to as logarithmic crowning) 1). When a
Lundberg curve is used, the distribution of contact**This paper is reprinted from the original paper (in Japanese) carried in the Proceedings of the Japan Society of
Mechanical Engineers Part C, Vol.72 (2006), pp.3022 - 3029**Elemental Technological R&D Center-140-

Logarithmic Profiles of Rollers in Roller Bearings and Optimization of the Profiles

designed easily and to give the technique of
optimizing it. To be more specific, we propose adesign technique, based on improving Johns-Gohar’sequation that will prevent edge stress due tomisalignment. A functional formula is introduced inwhich the degree of freedom is improved byintroducing three design parameters into JohnsGohar’s equation. Using this formula, one can set a
straight section of an arbitrary length in the contact
area. In addition, these parameters can be optimized
through the use of a mathematical optimization
method, which is shown below.

Outer ring

Flat regionzyx

Crowned region
Chamfer

Crown drop

Inner ring

Fig. 1 Schematic drawing of a crowned roller
in a cylindrical roller bearing

2. Johns-Gohar's logarithmic functional
formula and its improvement
Johns-Gohar improved Lundberg's formula,showing the following equation 2):

logarithmic shape prevents edge stress from occurring
even under the presence of misalignment, resulting ina long service life 3). However, he does not show thespecific shapes.
Takata et al. propose a crowning shape formed by
combining Lundberg’s logarithmic curve and circular
arcs that brings a long service life under the presenceof misalignment 4). This method has a drawback ofcomplicated calculation that express the amount ofcrowning corresponding to misalignment by means
circular arc, which is different than that of Lundberg’s
contact theory.
Paying attention to inner stress, Kamamoto et al.
present a functional formula that gives an optimum
shape 5). In other words, according to their theory,when Mises’ equivalent stress or Tresca’s equivalentstress appear near the contact area between the roller
and the race, the stress will become axially uniform,
which will minimize the damage the material
undergoes as well as elongate bearing life. However,the functional formula that Kamamoto et al. has given
to represent crowning curves does not take into
account the effect of misalignment.
Urata has proposed a crowning shape formed by
combining two or more circular arcs whose curvatures
diminish as they move from the center to the end of
the roller 6). Although he shows the crowningtechnique, edge stress does not appear in thecrowning shape even under the presence of
misalignment; the design method does not takemisalignment into consideration, with tolerable amountof misalignment not made clear.
In addition, all of the reports above do not givedetailed descriptions of the methods for optimum
crowning design. The purpose of this report is to
propose a logarithmic functional formula that can be

z（y）＝

2Q
πlE'

1

ln

2
1−
（1−0.3033b/a）（2y/l）

…
（1）

where
a
: 1/2 of the effective contact lengthb: Half width of contact
E'E: Equivalent Young's modulus ＝―――――
1−ν2

（

）

E : Young's modulus
ν: Poisson's ratiol: Effective contact length
Q : Loady: Position in the axial direction
z (y): Drop at the position in the axial direction y
However, when this equation is applied, carrying out
the calculation using the technique to be describedlater results in edge stress occasionally. This trend
becomes more remarkable under the presence of
misalignment. Providing a cylindrical roller or conicalroller with a straight section may be desirable formachining or functional reasons; however, using thisequation does not allow a straight section to be set up.
To solve this problem, three design parameters, K1,
Km, and zm, are introduced into equation (1) written it
into equation (2):

z（y）＝
×ln

2KlQ
πlE'
1
b
1− 1−0.3033Km a

（

-141-

）｛

y−
（a−ym）2 （2）l/2｝NTN TECHNICAL REVIEW No. 75（2007）

where ym denotes the length of the straight section
and is given by

ym＝（1−exp −zm

1
2

πlE'
2K1Q

1−0.3033Km b/a

Selecting these design parameters adequately
makes it possible to design a logarithmically crowned
shape that will not produce edge stress even if theroller is tilted.

） ………………（3）3. Method of calculating contact pressure

As described above, a straight section is set up on
the roller of a cylindrical roller bearing in some cases.
If equation (3) is used, the length of crowning ym canbe established by specifying K1, Km, and zm; however,directly giving ym as a design parameter is moreconvenient. Therefore, let us determine Km on thebasis of equation (3) by substituting Km into equation(2), which gives

Fig. 3 shows the schematic diagram of the crosssection at the contact area. Pressure occurs at thearea where two surfaces come in contact with eachother and no pressure occurs where there is nocontact. In other words, if the distance between thetwo surfaces is supposed to be h (x, y), the following
relationships are obtained:
h（x, y）＞0, p（x, y）＝0 Non-contact area

z（y）
＝
×ln

2KlQ

h
（x, y）＝0, p（x, y）＞0 Contact area

πl E'
1

）｛｝

（

When the elastic proximity amount between twobodies is given by ho (<0), the surface shape of asemi-infinite elastic body is given by g (x, y) and thedisplacement of the surface of the semi-infinite elastic
body is given by u (x, y) then distance between two
surfaces h (x, y) can be exoessed in the following
equation:y−（a−ym）2 （4）ymz πlE'
1− 1−exp − m
2KlQ

｛

｝

Further, defining K2 as shown below
ym = K2awhich gives the following equation:
z（y）
＝Aln

where,……
（6）1z
1− 1−exp − m
A

｛

A＝

（

）｝（

2
y−a
＋1
K2 a

）

h（x, y）＝h0＋g（x, y）＋u（x, y）…………………（7）…
（5）

The contact pressure p (η,ζ) and the displacementu (x, y) are in a relationship given by the following
equation

2K1Q
πl E'

u（x, y）＝

In this equation, the design parameters have thefollowing meanings:
K1 : Multiple of Q
K2 : Ratio of the crowning length to azm : Drop at the end of the effective contact length
K1 corresponds to the curvature of the crowned
section geometrically. Fig. 2 shows the portion thateach parameter represents and the corresponding
shape.

∞∞p（η,ζ）dηdζ
2
……（8）
2＋
πE' −∞ −∞ （x−η）（y−ζ）2∫∫In the above equation, (η,ζ) is the coordinate on
the xy-plane. In other words, equation (8) means thatthe displacement at (x, y) is affected by the contactpressure p at all points and that the magnitude of theeffect is inversely proportional to the distance between(x, y) and (η,ζ).
In addition, since the surface integral of the contact
Contact pressure
distribution

a

h
K1

u

zm

Roller profile

Deformed profileh (x, y)h0 (＜0)K2×a

Undeformed profile
g (x, y)Position of axial direction

Fig. 2 Logarithmic profile parametersgFig. 3 Contact deformation and pressure distribution-142-

Logarithmic Profiles of Rollers in Roller Bearings and Optimization of the Profiles

Even on Johns-Gohar's curve, edge stress having amaximum value of 3.75 GPa has occurred. On JohnsGohar's curve in the case of absence of crowning, thevalue of edge stress is reduced to 1/3. However, since
the edge stress value depends on the size of
calculation lattice, a simple comparison between
values is not allowed. However, since calculation
lattices of the same size are used in this case, adiscussion on relative sizes does not pose a problem.
On the other hand, in the case of crowningaccording to equation (2), under the conditions of Fig.
4, if the conditions are
K1＝1, K2＝1, zm＝12.7μm
The same values as Johns-Gohar's curve isachieved. However, if the conditions are
K1＝1.4, K2＝1, zm＝14μm
edge stress does not occur.
Setting K2 at 1 means providing the entire domain of
the effective contact length with crowning. If thecrowned domain is made into a domain half of "a" bysetting K2 at 0.5, the contact pressure distribution
exhibits a decrease at the middle of the contact area
and a slight increase at both ends. However, thisincrease in the contact pressure is not a spike-shaped
increase in the contact pressure resulting from thecontact of the intersection of the crowned part with thechamfer, thus differing from edge stress.
Fig. 6 shows the contact pressure distribution withthe roller tilted by 0.001 rad. On Johns-Gohar's curve,
the occurrence of edge stress is noted as in the casewhere the tilt angle is 0 rad.
Under the conditions of K1 = 1.4, K2 = 0.5, and zm =14 mm, edge stress did not occur, but under thecondition of the tilt angle of 0.001 rad, edge stressoccurred. Solve this problem by increasing zm from 14
mm to 17 mm, which results in the conditions of K1 =pressure is equal to the normal load Q, the following
equation holds:
∞

∞

−∞

−∞

∫∫ p（x, y）dx dy …………………………（9）Q＝The above simultaneous equations are solved
numerically by means of the iteration method. The
common iteration method has a drawback in that it isdifficult for long-wavelength error components toconverge against lattice spacing established in thespace for which calculation is to be carried out.
Furthermore, in the common calculation method, if thenumber of lattice points is set at n, the calculation ofequation (8) will take a time on the order of O (n2). To
handle this problem, the numerical calculation wasmade faster through the use of the multi-level method 7).
The multi-level method is a high-speed arithmetic
algorithm combining the multi-grid method that speeds
up the convergence calculation of simultaneous
equations by means of the iteration method and themulti-level multi-integration (MLMI) method that
executes the double integration of equation (8) at a
speed of O (n log n).

4. Contact pressure distribution on alogarithmically crowned roller
The contact between the cylindrical roller and theinner ring shown in Fig. 4 is considered. It is assumed
that the roller alone is crowned. It is also assumed thatboth the roller and the inner ring are made of bearing
steel. The calculation lattices are formed by dividing
the system into 256 sections in the axial direction andthe tangential direction into 32 sections.
Fig. 5 shows the relationship between the crowned
shape and the contact pressure distribution for a tilt
angle of 0 rad. When crowning is not provided, an
extremely large edge stress of 10.7 GPa occurred at
the end of the effective contact length of the roller.

No Crowning
Johns-Gohar

3.8

θ

d1

r

Q

Contact pressure GPaw
Roller4

d2

Inner ringw ＝10mmr ＝0.5mm
d1 ＝10mm

K1= 1.4, K2= 1.0, zm= 14μm
3.6

K1= 1.4, K2= 0.5, zm= 14μm

3.4
3.2
3
2.8

d2 ＝50mm
Q ＝9.8kN
θ ＝0 or 0.001rad

2.6
-5

-4

-3

-2

-1

0

1

2

3

4

5

Position of axial direction mm

Fig. 4 Schematic drawing of a roller and an inner ring used incontact pressure calculation

Fig. 5 Contact pressure distributions when the tilting
angle is 0 rad-143-

NTN TECHNICAL REVIEW No. 75（2007）

the initial value in terms of the convergence of asolution. The convergence of a solution means theconvergence solution itself and the amount ofcalculation required before the convergence solution
is obtained. It is important to determine a favorable
initial value in solving an optimization problem.
In this study, the initial value is determined in thefollowing method: Not only the contact pressure but
also the rolling fatigue life can be used as objectivefunctions. In the following description, the maximum
contact pressure Pmax is used as the objectivefunction.
(1) The limits of design parameters within which theinitial values are retrieved and the number ofvalues for each parameter within the limits are
specified.
(2) When numbers of values of K1, K2, and zm are l, m,and n, (l×m×n) combinations of design
parameters are defined. The calculation for thecontact pressure is carried out for all combinations
to determine Pmax.
(3) The combination yielding the smallest Pmax is
adopted as the initial value for optimization.

Johns-Gohar

4

K1= 1.4, K2= 0.5, zm= 14μm

3.8Contact pressure GPa

K1= 1.4, K2= 0.5, zm= 17μm
3.6
3.4
3.2
3
2.8
2.6
-5

-4

-3

-2

-1

0

1

2

3

4

5

Position of axial direction mm

Fig. 6 Contact pressure distributions when the tilting
angle is 0.001 rad

1.4, K2 = 0.5, and zm = 17 mm, and this prevents theedge stress from occurring.

5. Optimization of design parameters
5.1 Optimization algorithm

5. 2 Example of optimization

As described above, it is possible to decrease themaximum contact pressure by changing design
parameters. However, it is difficult to give the optimum
value by means of an analytical method. For thisreason, a technique of optimizing design parametersby means of a computer is considered.
Various algorithms including the conjugate gradientmethod, annealing method and genetic algorithm areproposed as optimization algorithms. In solving thisproblem, for example, if the maximum contactpressure or rolling fatigue life is selected as theobjective function it is impossible to analytically derive
its derivative function with numerical differentiationaccompanied by difficulty. For these reasons, we
adopt the Rosenbrock method 8), a direct retrieval
method. Its outline is as follows:(1) A directional vector is defined in the space ofvariables and the value of a variable is changed
along the directional vector.
(2) When a change in the variable improves theobjective function, the change is accepted. If theobjective function is degraded, the change isdiscarded.
(3) If convergence occurs after repeating steps (1) and(2), the operation is finished. If not, the directional
vector is modified and the operation goes back tostep (1).
Not only the Rosenbrock method but also otheroptimization techniques have strong dependence on

In the following, an example of optimization
calculation of the crowning of a cylindrical roller
bearing is shown. When only the roller is crowned, themaximum contact pressure between the roller and theinner ring is greater than that between the roller andthe outer ring. For this reason, when one considers
the optimization of crowning, considering the contact
between the roller and the inner ring suffices. It is
assumed that a crowned roller is tilted by half theamount of misalignment against the inner ring under amisalignment condition.
In the following, the calculation example uses thegeometrical shape and load condition shown in Fig. 4,
and it is assumed that both the roller and the innerring are made of bearing steel.

5. 2. 1 Case where the tilt angle is 0 rad
The optimization was carried out under theconditions shown in Table 1. Fig. 7 shows, in 3D
contour lines, the relationship between the design
parameters obtained by initial value retrieval and theTable 1 Optimizing condition No. 1-144-

Objective function

Maximum pressure

Roller tilting angle

0 rad

Initial value searching area of K1

1~3

Initial value searching area of K2

0~1

Initial value searching area of zm

5 ~ 20μm

Logarithmic Profiles of Rollers in Roller Bearings and Optimization of the Profiles

5

3.4-3.5

5

zm μm

K1

1.0

18.5

（a）K2=0.0

5

zm μm

Pmax GPa

3.7-3.8
3.6-3.7
3.5-3.6
3.4-3.5
3.3-3.4
3.2-3.3

5

K1

zm μm

5

2.8

zm μm

3.2-3.3

K1

1.0

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1

Pmax GPa

3.7-3.8
3.6-3.7
3.5-3.6
3.4-3.5
3.3-3.4
3.2-3.3

5zm μm（e）K2=0.8

1.6
18.5

1.0

K1

（f）K2=1.0

Fig. 7 Design parameters and maximum pressure under the optimizing condition No.1

-145-

3.7-3.8
3.6-3.7
3.5-3.6
3.4-3.5
3.3-3.4
3.2-3.3
3.1-3.2

2.2

14

K1

3.8-3.9

2.8
9.5

3.1-3.2

2.2
1.0

3.4-3.5

3.1-3.2

3.8-3.9

1.6
18.5

3.5-3.6（d）K2=0.6

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1

14

3.6-3.7

1.6
18.5（c）K2=0.40

3.7-3.8

3.3-3.4

2.2

14

1.0

3.8-3.9

2.8
9.5

3.1-3.2

2.2

9.5

K1

1.0

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1

1.6
18.5

3.1-3.2

3.8-3.9

2.8
14

3.4-3.5（b）K2=0.2

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1
9.5

3.5-3.6

1.6

Pmax GPa

18.5

3.6-3.7

3.2-3.3

2.2

14

1.6

3.7-3.8

3.3-3.4

2.8
9.5

3.1-3.2

Pmax GPa

3.5-3.6

3.8-3.9

Pmax GPazm μm

3.6-3.7

3.2-3.3

2.2

14

3.7-3.8

3.3-3.4

2.8
9.5

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1

3.8-3.9

Pmax GPa

3.9
3.8
3.7
3.6
3.5
3.4
3.3
3.2
3.1

NTN TECHNICAL REVIEW No. 75（2007）

Table 2 Optimizing condition No. 2

maximum contact pressure. In Figs.7 (a) to 7 (f), K2 ischanged in increments of 0.2 between 0 and 1. The
individual graphs show the relationship among K1, zm,
and Pmax for different K2 values.
The best initial values are the following ones asshown in Fig. 7 (e).
K10＝1.4
K20＝0.8
zm ＝12.5μm
As a result of the optimization based on the
Rosenbrock method, the optimum design parameters
are obtained as shown below:
K1＝1.295
K2＝0.879
zm ＝12.684μm
When a crowning operation is carried out using the
optimum design parameters, the contact pressure
distribution is almost uniform along the axis as shownin Fig. 8.

Maximum pressure

Roller tilting angle

0.001 radInitial value searching area of K1

2~4

Initial value searching area of zm

10 ~ 30μm

Contact pressure GPa

4.5
4
3.5
3
2.5
2
1.5
1
0.5
0
-5

-4

-3

-2 -1

0

1

2

3

4

5

Position of axial direction mm

Fig. 9 Pressure distribution of the optimized profile under thecondition No.2 ( K1 = 2.779, K2 = 0.5, zm = 16.253 mm )4.5

Contact pressure GPa

Objective function4
3.5

causes the contact pressure distribution in the middle
to become concave.
In this way, this technique allows the optimum
logarithmic crowning to be designed in cases where atilt or straight section exists.

3
2.5
2
1.5
1
0.5
0
-5

-4

-3

-2 -1

0

1

2

3

4

5. 2. 3 Case where life is taken as the objectivefunction5

Position of axial direction mm

The rolling fatigue life of the roller can be used as
the objective function for the optimization. The life iscalculated using Harris' method 9). Table 3 shows theconditions for optimization.
The best initial values are:K10＝1.2
K20＝0.9
zm ＝11μm
The following values are obtained as the best
possible design parameters through optimization:
K1＝1.065
K2＝0.982
zm ＝10.799μm
Under the optimization conditions adopted here,edge stress is allowed to occur. The contact pressure
distribution then is as shown in Fig. 10 with edge
stress occurring at both ends.

Fig. 8 Pressure distribution of the optimized profile under thecondition No.1 ( K1 = 1.295, K2 = 0.879, zm = 12.684 mm )5. 2. 2 Case where the tilt angle is 0.001 rad
We consider a case where the roller was tilted anda straight section was provided. In other words, K2 isfixed at 0.5 to be excluded from the objects to be
optimized. The optimization conditions for this setting
are shown in Table 2.
The best initial values are:K10＝2.8
zm 0＝16μm
The best possible design parameter values are
obtained through optimization.
K1＝2.779
zm 0＝16.253μm
The contact pressure distribution then is as shownin Fig. 9.
The presence of a tilt causes the maximum contactpressure to occur at a negative position in the axial
direction, and the provision of a straight section

5. 2. 4 Shape formed as a result of logarithmiccrowning
Fig. 11 shows the shape formed as a result of
logarithmic crowning optimized under the conditions
described above. Comparing "optimization conditions-146-

Logarithmic Profiles of Rollers in Roller Bearings and Optimization of the Profiles

Table 3 Optimizing condition No. 3
Objective function

Rolling fatigue life

Roller tilting angle

0 rad

Initial value searching area of K1

1~3

Initial value searching area of K2

0~1

Initial value searching area of zm

5 ~ 20μm

pressure is, the shorter the life is; however, an
increase in local contact pressure like edge stress has
only a small effect according to Harris' method of
calculating the life. For these reasons, whether edge
stress occurs or not depends on which of the
maximum contact pressure and life is adopted as theobjective function for the optimization regardless ofthe similarity of the shape of rollers.

Contact pressure GPa

4.5
4
3.5

6. Conclusion

3

We made a proposition regarding a method ofdetermining design parameters in the design oflogarithmic crowning for roller bearings, by introducing
the design parameters into a logarithmic crowning
equation and using a numerical optimization method.
Specifically, we improved Johns-Gohar's functional
formula to present a logarithmic crowning equation
with an improved degree of design freedom. The
design parameters introduced are the three that
follow:
K1: Scale factor for design load, which affects thecurvature of a crowned part in terms of geometry
K2: Ratio of the crowned part to the length from thehome position to the chamfered partzm: Drop at the end of the effective contact length
K2 may be defined at the time of designing the basic
shape; in such a case, K1 and zm only are optimized.
The optimization condition is either the maximum
contact pressure at its lowest or the maximum life. The
Rosenbrock method is used as the optimization
algorithm. By numerically optimizing these design
parameters, the optimum logarithmic crowning for aroller bearing can be obtained even under thepresence of misalignment.
In the following, the features of this technique are

2.5
2
1.5
1
0.5
0
-5

-4

-3

-2 -1

0

1

2

3

4

5

Position of axial direction mm

Fig. 10 Pressure distribution of the optimized profile under thecondition No.3 ( K1 =1.065, K2 = 0.982, zm = 10.799 mm )1 (objective function: maximum contact pressure; tilt
angle: 0 rad)" and "optimization conditions 3 (objectivefunction: life; tilt angle: 0 rad)" reveals little difference
between the shapes. However, when the life is taken
as the objective function, edge stress occurs asshown in Fig. 10. Fig. 7 shows that, with K2 fixed at avalue, the maximum contact pressure increases
abruptly when K1 or zm becomes slightly smaller thanthe optimum point for the case where the maximum
contact pressure is taken as the objective function.
This increase in the maximum contact pressure is
generated by the occurrence of edge stress.
Qualitatively speaking, the greater the contact

20

Optimization condition No.1
(Objective function : maximum pressure,tilting angle 0 rad)18

Roller profile μm

1614

Optimization condition No.2
(Objective function : maximum pressure,tilting angle 0.001 rad)12
10

Optimization condition No.3
(Objective function : rolling fatigue life,
tilting angle 0 rad)

8
6
4
2
0
-5

-4

-3

-2

-1

0

1

2

3

4

5

Position of axial direction mm

Fig. 11 Optimized roller profiles under the conditions No. 1 to 3

-147-

NTN TECHNICAL REVIEW No. 75（2007）

summarized:(1) In considering misalignment in the design oflogarithmic crowning, the effect of themisalignment is treated independently of thelogarithmic function. Contrary to this, thistechnique treats the effect as a design parameterand incorporates it into the logarithmic equation.
(2) With the three design parameters corresponding to
the crowning shape, the features of a crowned
shape can be expressed by the parameter values
only.
(3) Incorporating the straight section into a design
parameter enables the straight section to be
combined with logarithmic crowning easily andwith a greater degree of freedom.
(4) A new design technique is proposed ofincorporating the numerical value optimization
technique into logarithmic crowning design.

References
1) Lundberg, G., Elastic Contact Between Two SemiInfinite Bodies, Forschung auf den Gebiete desIngenieurwesen, 5(1939), pp.201-211. (in German)
2) Johns, P. M. and Gohar, R., Roller bearings underradial and eccentric loads, Tribology International,
14(1981), pp.131-136.
3) Reusner, H., The logarithmic roller profile ― the key tosuperior performance of cylindrical and taper roller
bearings, Ball Bearing Journal, 230(1987), pp.2-10.
4) Takata, H. et al., Experimental Study of Fatigue Life ofProfiled Roller Bearings, NSK Technical Journal,653(1992), pp.1-7. (in Japanese)5) Kamamoto, S. et al., Research on Crowning Profile toObtain The Maximum Load Carrying Capacity for
Roller Bearings, KOYO Engineering Journal,
159(2001), pp.44-51.
6) Urata, S., Investigation of Optimum Crowing Profile of
Cylindrical Roller Bearings Part 2, FUJIKOSHIEngineering Review, 56(2000), pp.14-23. (inJapanese)
7) Venner, C. H. and Lubrecht, A. A., Tribology Series, 37
Multilevel Methods in Lubrication, (2000), Elsevier
Science B. V.
8) Bazarra, M. S. et al., Nonlinear Programming, (1993),
p.291, John Wiley & Sons.
9) Harris, T. A., Rolling Bearing Analysis, Forth Edition,(2000), pp.728-729, John Wiley & Sons.

Photos of authors

Hiroki FUJIWARA

Tatsuo KAWASE

Elemental Technological
R&D Center

Elemental Technological
R&D Center-148-

NTN TECHNICAL REVIEW No. 75（2007）

Presentation of award-winning articles
Japan Society for Design Engineering’s 2005 Most Interesting Reading Prize

Run-out Analysis for Rolling Element Bearing

Tomoya SAKAGUCHI*

1. Introduction

2. Summary of the commentaryThis is a commentary on rotational accuracy
analysis of rolling bearings. The commentary
introduces an example of study on effects of shapeerrors in balls and inner and outer rings and of angularintervals between balls on rotational deflections in thethree axial directions of the inner ring; the study isbased on quasi-static analysis that takes into
consideration translational displacement of balls andinner rings on the radial plane in ball bearings. The
commentary also presents the result of study on
rotational accuracy.
This commentary was carried in Journal of Japan
Society for Design Engineering, Vol.40, No.10, 2005.
The Society commended the commentary as the mostinteresting among the articles other than treatises
carried in the journal in the year. The analyticaltechnique introduced in the article is as described in
our Review No.69.

To analyze effects of errors in geometries of bearing
components and of angular intervals of balls on
rotational accuracy of ball bearings, a simulation
program was developed in which the contact force
between a rolling body and the raceway surface ismodeled by a linear spring on a radial plane (Figs. 1
and 2) and the angular position of balls can bearbitrarily set up.
Through experimental verification, the validity of thissimulation was confirmed. In addition, the result of astudy on the relationship between errors in the shapeof components and the rotational accuracy revealed
that the rotational accuracy deteriorates when
raceway surface deviations from a perfect circle withwaviness orders Z and Z±1 exist, where Z denotes
the number of balls.
When no misalignment exists under pure axial load,
it is thought that the practically sufficient rotational
accuracy of a ball bearing can be estimated. It is
anticipated that, in future, technologies for analyzing
rotational accuracy under radial loading will evolvewith the aim of reducing noise of increasingly fastrotating machines.

Bearing under consideration

Analytical modelFig. 1 Sections of an actual ball bearing with an axial
load and the simplified model

Photo of author※An arrow denotes a linear spring

Ob,1

Ob,2
Oi
Tomoya SAKAGUCHI

Oe
Fig. 2 Schematic of linear springs on radial plane of abearingElemental Technological
R&D Center-149-

NTN TECHNICAL REVIEW No. 75（2007）

Presentation of award-winning articles
The Fourth Monozukuri Parts Grand Prix, Incentive Award

Improving Rolling Contact Fatigue Life of Bearing Steels
Through Grain Refinement "Development of the FA bearing"
Chikara OHKI*

exhibited an improvement in breaking stress andindentation forming contact pressure (the maximum
contact pressure that causes the residual indentation
depth to reach 1/10000 of the diameter of the rolling
element when it is pressed to the surface), which also
increased the rolling fatigue life. Table 1 shows an
overview of the performance.
It is certain that the development of structural metalmaterials will advance toward grain refinement, andthis technology is expected to be unfolded broadly. By
considering future world trends, it can be said that it is
not desirable to increase steel strength by adding alloy
elements or through inefficient manufacturing
processes. Our goal is to reduce the environmental
impact and conserve energy through widespread useof FA-treated bearings.

Summary
Improvement in strength for iron-based materials
through reinforcement by means of grain refinement
according to Hall-Petch’s law is well known, notably inthe ferrite phase. It has already been confirmed thatthe same principle applies to austenite and martensite
phases. In ferrite and austenite-based materials,
grain refinement is relatively easy to perform withsuccessful commercialization in the field of controlled
rolling. On the other hand, it is difficult to put
martensite-based materials like bearing materials,
which need a quenching process in their heat
treatment, to grain refinement. As for specific methods
for grain refinement, two examples are the Grange
method that is based on repeated quenching and theausforming method, but there are also several others
that can be listed. However, all of these processes
need excessive man-hours, which is not a practicalindustry practice.
The FA treatment developed this time does notneed a large amount of man-hours. Through the
improvement of heat treatment the results areenhanced grain refinement, enhanced solid solution of
nitrogen, and enhanced precipitation. Fig. 1 showsprior-austenite grains subjected to the standard heat
treatment and the FA treatment. The steel type is
SUJ2. Compared with the grain boundary after
standard heat treatment, the grain boundary of FA
treatment shows that the average grain size of prioraustenite grains has been reduced to about 5mm,
which is half of the prior size. The FA-treated product

Table 1 A summary of characteristics
Ratio of a characteristic of an
FA-treated product to that of aconventional carbo-nitridingtreated one*1
Breaking stress

1.2

Indentation forming contact pressure*2

1.1

Change in dimensions due to aging*3

0.7

Lubrication life with mixed foreign bodies*4

2*1 Referred to a characteristic of a conventional carbo-nitriding-treated product set
at 1*2 Maximum contact pressure that causes the residual indentation depth to reach
1/10000 of the diameter of the rolling element when it is pressedto the surface*3 A product showing a smaller value after having been held at 120˚C for 2500
hours is a better one.
*4 Tapered bearing 30206; contact pressure: 2.5 GPa
Foreign bodies: One g/L of gas-atomized HV800 (50μ or smaller: 100-180μm =9:1) mixed in the lubricant

Photo of author

0.05mmFA-treated product0.05mmChikara OHKI

Standard-heat-treated productElemental Technological
R&D Center

Fig. 1 Prior austenite grain boundary-150-

NTN TECHNICAL REVIEW No. 75（2007）

Presentation of award-winning articles
Outstanding Prize, the 12th Advanced Display of the Year 2007

（NRS-3000 Series））
LCD Color Filter Repair System（
Akihiro YAMANAKA*

1. Introduction
In 1990, NTN commercialized devices to repair bridging
defects on liquid crystal TFT (thin-film transistor) electrodes
by laser-cutting them, making its way into the field of defectrepair devices for FPDs (flat panel displays). Since then,NTN has accumulated a number of track records and many
years of experience through the development of functions ofapplying ink on micro-domains stably and other functions ofrepairing defects.
Recently, NTN was given the ADY Outstanding Prize inrecognition of its contribution to energy conservation in FPD
manufacturing processes. NTN's contribution resulted fromits efforts to develop an ink application mechanism thatenables the quality of repair to be improved and repair tacttime to be reduced and also from its efforts develop afunction to automatically repair defects produced on color
filters for liquid crystal display use.
This award is given to commend, by section of
specialization, outstanding FPD-related devices sold in thepast year. The color filter repair device was commended in
the inspection, repair, and measurement section. Photoshows the appearance of the device. (Approximate
dimensions of the device: 3 m in width×4 m in depth×3 m
in height)

grinding them partially.
Particle defect
Clear defectEquipped with an ink
application mechanism thatwas greatly improved series
NRS-3000, commended
this time--shortens theOpaque defectrepair tact time to one thirdof the time a conventional
Fig. Defects in a CF panel
NTN device takes. In
addition, to achieve an improvement in the quality of repair,the tip of the application needle, used to apply ink, wasspecially processed to permit the quantity of ink application
to be controlled in a range between several pico liters andseveral tens of pico liters (a pico liter is a trillionth of a liter)
depending on the time during which a substrate remains incontact with the application needle. This enables us torespond to the recent requirement for ever increasing
resolution for liquid crystal displays.
In addition, the repair process was considered to bedifficult to automate and operators checked defects todetermine repair conditions. However, NTN developed aunique image processing algorithm, being the first toachieve the automatization of the repair process in theindustry. This innovation has made it possible to remove an
operator needed for each device in the conventional
process.

3. Summary
In addition to the color filter repair device presented in thiscommentary, NTN has developed and expanded a numberof repair devices for FPDs including defect repair devices forplasma displays. Recently, new displays such as SEDs
(surface-conduction electron-emitter displays), organic EL(organic electroluminescence) displays, and electronic paper
are being developed rapidly and put into mass production.
NTN will continue its effort in developing devices that willcontribute to the progress of such new displays.

Photo Appearance of the device

2. Outline of the color filter repair device
In the manufacturing process of color filters for liquid
crystal displays, clear defects (color dropout), opaquedefects (mixed colors), and particle defects (sticking of dust )occur occasionally. A color filter repair device is intended for
repairing such defects.
Clear defects are repaired by coating clear portions withink having the same color as the original pixel color. Opaquedefects are repaired by first cutting defective portions with
YAG laser to change them into clear defects and thencoating them with ink. Particle defects are repaired by first
cutting defective portions with laser in the same way as inrepairing opaque defects and then covering them with ink or

Photo of author

Akihiro YAMANAKA
Product Engineering Department
Precision Equipment Division-151-

NTN TECHNICAL REVIEW No.75（2007）[ New Products Information ]Hub bearing equippedwith high resolution rotational sensor
Contributing to better vehicle safety control

Features

Construction(1) High resolution rotational signal (40times higher relative to conventional
types; magnetic encoder is the same asconventional types)(2) A & B phase pulse output allows
detection of rotational direction(3) External dimensions are the same asconventional hub bearings and can beapplied to both driven and non-driven
wheelsMagnetized
encoder

Applications¡Hub bearings for passenger vehicle use-152-

SNR madehigh resolution
rotational sensor

New Products Information

Environment-friendly roller bearing
Using biodegradable materials to make the cage, seal and grease, the
new roller bearing reduces environmental loading after being discarded

Features

Construction¡The cage and seal are made up of apolyester-based biodegradable resin
having improved strength and heat
resistance.

(1) The biodegradability of the cage andseal is over 60%.
[A product eligible for a Green Pla Mark](2) The biodegradability of the grease isover 60%.
[A product eligible for an Eco Mark]¡The grease is made up of a base oilcomposed of an ester-based synthetic oil
having a 60% biodegradability.

(3) Working temperature limits: –30 to 70˚C(4) Allowable number of revolutions: dmn
value: 350,000(about the same as that of existing
general-purpose products)※ Green Pla Mark: An official certification mark given to a biodegradable plastic product.
※ Eco Mark: An official certification mark given to a biodegradable material (other thanplastics) or a commodity using such material.

-153-

NTN TECHNICAL REVIEW No.75（2007）

World’s smallest tapered roller bearing
A tapered roller bearing with a bearing inside diameter of five mm achieves alonger life and higher rigidity than similar ball bearings

Features

Construction(1) Tapered roller bearing with the world’s
smallest diameterFive-mm inside diameter φ17-mm
outside diameter φ10-mm width(2) Longer life
Fourteen-time longer life than that ofsimilar ball bearings(3) Higher rigidity
Five-time greater axial rigidity than thatof similar ball bearings[Cross-section of
the developed product]

φ5

φ19

φ5

φ17

8

Comparison of axial rigidity
Axial displacement μm

610[Cross-section ofa similar ball bearing]

30
25
20

Applications

15635
CR-501

10¡Reducers, medical equipment, robot
joints, and the like

5
0

0

1

2

3

4

5

Axial load kgf
Axial rigidity: Five times(than that of similarly-sized ball bearings)-154-

New Products Information

MQCJ lubrication angular contact ball bearing
The world’s highest speed angular contact ball bearing has been realized
through the adoption of a new jet lubrication scheme equipped with inner ringcooling and a mechanism for best minimum quantity lubrication for the raceway

Features

Construction (means of realization)(1) High-speed operation at the world’s
highest speed• Constant pressure pre-load:
dmn value: 5 million

1. The adoption of inner ring cooling scoop
grooves and special nozzle spacers hasrealized inner ring cooling and best
lubrication.
2. Optimization of the bearing interior design has
realized ultra-high speed and high rigidity.
3. Special inner/outer ring materials excel indurability under high-speed, high-contactpressure operation.

(90% increase in the conventional air/oil lubrication ratio)• Constant position pre-load:
dmn value: 3.6 million(40% increase in the conventional air/oil lubrication ratio)(2) Power loss: on the same level as that of
air/oil lubrication(3) Simplification of the oil supply /discharge system
The outer cylinder cooling oil serves
also as the bearing lubrication oil.

Lubricant [1 to 2.5 L/(min·bearing)]

Best minimum
quantity lubricant

Applications

Inner ringjet cooling¡Ultra-high-speed main spindle for machine
tools (machining centers for die machining
and aluminum machining)*MQCJ stands for Minimum Quantity and Cooling Jet

Discharging the inner ring cooling oil and bearing lubricant-155-

NTN TECHNICAL REVIEW No.75（2007）

Angular contact ball bearing with ultra-high-speed seal
The new grease SE-1 and optimization of internal design have realized
ultra-high-speed operation at the world’s highest level

Features

Construction
1. Optimization of internal design⇒High speed and low heat generation
2. Adoption of the newly developed grease
SE-1⇒High speed and long life
3. Adoption of a specially developedmaterial for the race
⇒Improved seizure resistance(1) Angular contact ball bearing with seal
having high-speed rotation performanceat the world’s highest level
dmn value: 1.7 million; 20% improvementcompared to the performance of
conventional similar products(2) The bearing washing process before
mounting and the grease filling process
are not necessary.
(3) External lubricant feeders are notnecessary.

Applications¡Machining tool main spindle (machining
centers and tapping machines)
Cross-section of an angular contact ball bearing
with ultra-high-speed seal

-156-

New Products Information

New high-speed specification bearing unit for rolling stocks(New RCT bearing)
The bearing maintenance interval is lengthened more than twice

Features

Construction(1) Special-resin-made cage
Longer lubrication life as a result of the
adoption of a special, anti-shock resin
and the suppression of abrasion loss
(2) Spacer with rubber lip
Suppression of fretting between the
inner ring and the back cover due todynamic deflection of a shaft(3) New shape seal
Longer lubrication life as a result of
reduced temperature rise

New shape seal

Spacerwith rubber lip

Applications

Brass plate
Resin-made cage

¡Shaft support bearings for rolling stock
use

-157-

Rubber lip

NTN TECHNICAL REVIEW No.75（2007）

Office equipment paper-reversing unit
Smooth paper feed without a reversing motor in the double-side printing
mechanism for printing and copying machines

Features

Construction¡The unit consists of three gears, namely, an
input gear, an output gear, and a control gear,with the input gear always rotating in one(normal) direction.
¡When the control gear is not braked (free),both the control and the output gear rotate in
the normal direction (the same direction asthat of the input gear).
¡On the other hand, when the control gear isbraked to a halt (locked), the output gearrotates in the direction opposite to that of the
input gear.

(1) The output gear can be rotated bothclockwise and counterclockwise at thesame speed with that of the input gear.
(2) A dedicated reversing motor is notneeded.

Applications¡Paper feed direction switchovermechanisms for copying machines andprinters capable of both side printing

Normal rotation input
Control gear

Free

Locked-158-

Input gear

Output gear

Normalrotation

Reverserotation

New Products Information

Color filter automatic repair machine
Automatic repairing of defects in color filters for flat panel displays
improves labor efficiency, saving labor

Features

Defect repair methods(1) NTN’s unique image processing
technology is used to locate a defect,
automatically remove (laser-cutting) the
defect and coat it according to its typeand color.
(2) Automatic repairing improves labor
efficiency, saving labor.

Coating position

Clear defect repairing(The circle denotes the result of theautomatic coating location calculation.)

Applications

Laser cut position¡Color filter defect repair(Opaque defects, clear defects, mixed colors)

Opaque defect repairing(The square denotes the result of thelaser cutting location calculation.)-159-

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NTN CORP. URL http://www.ntn.co.jp

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SALES NETWORKNTN BEARING CORP. OF AMERICANTN BEARING-THAILAND CO., LTD.

Head Office / 1600 E. Bishop Court, P.O. Box 7604, Mount Prospect, IL 60056-7604,
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Phone: +1-770-448-4710 Fax: +1-770-448-6969
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Phone: +1-817-329-1818 Fax: +1-817-329-4711
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IL 60056-7604, U.S.A.
Phone: +1-847-699-4060 Fax: +1-847-294-1364
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Phone: +1-248-324-4700 Fax: +1-248-324-1103

Head Office / 12th Floor, Panjathani Tower, 127/15 Nonsee Road, Chongnonsee
Yannawa, Bangkok 10120, Thailand
Phone: +66-2-681-0401 Fax: +66-2-681-0408
Khon Kaen Branch / 189-191 Ruen Rom Road Tambon Nai-Muang, Amphur
Muang, Kohn Kaen, 40000, Thailand
Phone: +66-43-223679 Fax: +66-43-223061
Haad Yai Branch / 198-198/1 Nipat U-Thid 2 Road, Amphur Haad Yai, Songkhla,
90110, Thailand
Phone: +66-74-236568 Fax: +66-74-231520
Bangna Sales Office / 35/35 Bangna-Trad Road, KM11 Bangplee, Samutprakarn,
10540, Thailand
Phone: +66-2-7501732 Fax: +66-2-7501731
Chiangmai Sales Office / 94. 94/1 Chayapoom Road, Tumbolsripoom, Amphur
Muang, Chiang Mai 50200, Thailand
Phone: +66-53-874328 Fax: +66-53-874330

NTN BEARING CORP. OF CANADA LTD.
Head Office / 305 Courtneypark Drive West, Mississauga, Ontario, L5W 1Y2, Canada
Phone: +1-905-564-2700 Fax: +1-905-564-7749
Vancouver Branch / 201-669 Ridley Place Annacis Island Delta, B.C., V3M 2Y9 Canada
Phone: +1-604-517-1777 Fax: +1-604-517-1794
Edmonton Branch / 4608-97th Street, Edmonton, Alberta T6E 5N9, Canada
Phone: +1-780-435-6200 Fax: +1-780-435-3600
Toronto Branch / 305 Courtneypark Drive West, Mississauga, Ontario L5W 1Y4,
CanadaPhone: +1-905-564-9600 Fax: +1-905-564-9609
Montreal Branch / 4973 Levy Street, Ville, St-Laurent, Quebec, H4R 2N9, Canada
Phone: +1-514-333-8054 Fax: +1-514-333-1078

NTN WÄLZLAGER (EUROPA) G.m.b.H.
Head Office / Nordrhein-Westfalen Branch /Max-Planck-Strasse 23, 40699 Erkrath, F.R.Germany
Phone: +49-211-2508-0 Fax: +49-211-2508400
Stuttgart Branch / Schurwaldstrasse 13, 73765 Neuhausen, F.R.Germany
Phone: +49-7158-17040 Fax: +49-7158-170460
München Branch / Geretsrieder Strasse 10A, 81379 Munchen, F.R. Germany
Phone: +49-89-7488630 Fax: +49-89-786382
Hamburg Branch / Barkhausenweg 7, 22339 Hamburg, F.R. Germany
Phone: +49-40-536962-0 Fax: +49-40-53696215
Italy Branch / Via Maestri del Lavoro, 3/A, 40138 Bologna, Italy
Phone: +39-051-535174 Fax: +39-051-538492

NTN BEARINGS (UK) LTD.
Wellington Crescent, Fradley Park, Lichfield, Staffordshire, WS13 8RZ, U.K.
Phone: +44-1543-445000 Fax: +44-1543-445035

NTN BEARING-MALAYSIA SDN. BHD.
Head Office / No.2, Jalan Arkitek U 1/22, Hicom Glenmarie Industrial
Park, 40150 Shah Alam, Selangor, Malaysia
Phone: +60-3-55696088 Fax: +60-3-55690200
Butterworth Branch / 4700, Jalan Permatang Pauh, 13400 Butterworth, Malaysia
Phone: +60-4-3328312 Fax: +60-4-3324407
Ipoh Branch Office /65, Medan Kidd, Kinta Mansion, 30200 Ipoh, Malaysia
Phone: +60-5-2547743 Fax: +60-5-2538077
Kuantan Branch / B-72, Ground Floor, Jalan Beserah 25300 Kuantan, Malaysia
Phone: +60-9-5141132 Fax: +60-9-5141164
Johor Bahru Branch / 51 Jalan, Sri Bahagia 5, Taman Sri Bahagia, Tampoi, 81200
Johor Bahru, Malaysia
Phone: +60-7-2364929 Fax: +60-7-2370897

NTN-CBC (AUSTRALIA) PTY. LTD.
3, The Crescent, Kingsgrove, NSW 2008, LOCKED BAG 1800, Kingsgrove 1480.
NSW Australia
Phone: +61-2-9502-1833 Fax: +62-2-9502-4013

NTN DE MEXICO, S.A.
Guadalajara Office / Calle 22 No.2465, Esq. Calle 3, Zona Industrial, C.P. 44940
Guadalajara, Jalisco, México
Phone: +52-33-3145-1490 Fax: +52-33-3145-1594
México Office / Calle Emilio Cardenas No.158, Fracc.Industrial San Nicolas
C.P.54030 Tlalnepantla, Estado De Mexico
Phone: +52-55-5565-5562 Fax: +52-55-5565-8638

NTN SUDAMERICANA, S.A.
World Trade Center Panama
Calle 53 Este, Urbanización Marbella Piso NO.16, Oficina 1601
Apartado Postal 832-0487, Panamá, Rep.de Panamá
Phone: +507-269-4777 Fax: +507-264-5592

NTN FRANCE S.A.
Head Office / Z.I.Sablière BP 338 Schweighouse Sur Moder 67507 Haguenau Cedex, France
Phone: +33-3-88-53-22-22 Fax: +33-3-88-73-46-95
Lyon Branch / Parc Technologique 2 Pace Berthe Morisot 69792 Saint-Priest
cedex, France
Phone: +33-4-72-04-0044 Fax: +33-4-72-04-4456
Paris Branch / Boulevard De Beaubourg B.P. 27 Emerainviller, 77313 Marne-lavallee Cedex 2, France
Phone: +33-1-64-80-4747 Fax: +33-1-64-80-4778

NTN BEARING-SINGAPORE (PTE) LTD.
Head Office / No.9 Clementi Loop Singapore 129812
Phone: + 65-64698066 Fax: +65-64695400
Philippine Representative Office / Unit 1002 Philippine Axa Life Centre Condominium
Corporation
Sen. Gil Puyat Ave Corner Tindalo Street Makati City, Philippines
Phone: +63-2-759-4407 Fax: +63-2-759-4409
India Representative Office / 805, International Trade Tower, Nehru Place, New Delhi
110019, India
Phone: +91-11-51513234 Fax: +91-11-51513236

NTN DO BRASIL LTDA.
Av. Moema, 94-9˚ Andar-conj, 92a94 CEP 04077-020-Indianópolis-São Paulo-SP,-Brasil
Phone: +55-11-5051-0600 Fax: +55-11-5051-2807

NTN KOREA CO., LTD.
Head Office / 10th Floor, Press Center, 25, Taepyeong-Ro 1-GA, Jung-Gu, Seoul
100-745, Korea
Phone: +82-2-720-3665 Fax: +82-2-720-3669
Pusan Branch / Rm.707, 7th Floor, Daerim Bldg., 341-5 Bujeon 1-Dong, BusanjinGu, Busan, Korea, 614-843
Phone: +82-51-811-1351 Fax: +82-51-811-1353

NOTE : The appearance and specifications may be changed without prior notice
if required to improve performance. Although care has been taken toassure the accuracy of the data compiled in this catalog, NTN does notassume any liability to any company or person for errors or omissions.
Printed in Japan

NTN TECHNICAL REVIEW No.75
Printed and published in October 31, 2008
EditorKenji OKADA

Publisher Keiji OOHASHI
Published by NTN Corporation
3-17, 1-chome, kyomachibori, Nishi-ku, Osaka,J apan
Printed by
NISSHA Co., LtdNo part of this publication may be reprinted or reproduced
without the prior permission of the copyright owner.