Special Issue;
Automotive Products
October 2005

ISSN 0915-0528 OSAKA,JAPANGeneral Research & Development Center Established in an Effort toEnhance New Product Development Capabilities

In September 2005, NTN's General Research & Development Center was completed in Iwata City, Shizuoka
Prefecture, Japan, as the keystone of NTN's R&D network of four R&D bases globally. Having a total floor area of
16,775 m2, the Center's five-story building possesses a seismically isolated steel structure that incorporates Super
Sliding Bearings manufactured by NTN Engineering Plastics Corporation. The Center Building also has an engine
driven power generator so that NTN's global research and development system will not be affected by the Tokai
Earthquake, a strong earthquake that is expected to occur in the near future in this region.
NTN considers the General R&D Center to be the core of NTN's R&D efforts to create new products andtechnologies for the world. The researchers at the Center work around the clock in cooperation with local engineers
and technicians to cope with technical assistance requests from customers around the world. Through this work theCenter seeks to accelerate development of new products and ensure swift technical support for customers around theglobe. Furthermore, the Center will help NTN develop new products in cutting-edge industrial fields, including next
generation automobiles, environmental and energy applications, medicines and robotics. In particular, NTN has been
augmenting the team of engineers that specializes in electronics applications and research in order to develop
promising new products.
On the fifth floor of the Center, an exhibition hall displays NTN's newest products and technologies. We hope that
visiting customers will fully appreciate NTN's capabilities for proposing and developing novel products andrecognizing NTN as a valuable engineering partner.

[Features of the General Research & Development Center]
Location: On site at NTN Iwata Works (1578 Higashi Kaizuka, Iwata-shi, Shizuoka Prefecture, Japan)
Total floor area: 5 stories, 16,775 m2
Structure: Seismically resistant building that incorporates elastic Super Sliding Bearings (SSB) manufactured by NTN
Engineering Plastics Corporation
Equipment: Emergency relief equipment including an engine driven power generator and a well-water utilization system;
roof garden; eco air-conditioning system

CONTENTS
Kenji OKADA

Preface

Eco-consciousness

Contribution

Technology on Environmental Protection for The Next Decade
Nobuo IWAI

The Reduction of Hazardous Substances
Kiyoshi NAKANISHI and Masakazu HIRATA

Application of Topology Optimization and Shape Optimization for Development of Hub-Bearing Lightening
Haruo NAGATANI and Tsuyoshi NIWA

Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing
Kazuyoshi HARADA and Tomoya SAKAGUCHIHigh Capacity Tapered Roller Bearings
Takashi TSUJIMOTO and Jiro MOCHIZUKI

Study of Long-Life Thrust Needle Roller Bearings Used in Low Viscosity Lubrication Conditions
Hiroki FUJIWARA and Kenji TAMADA

Double-Row Thrust Needle Roller Bearings for Automotive Air Conditioners and Automatic Transmissions
Kosuke OBAYASHI

Development of In-Wheel Type Integrated Motor Axle Unit
Minoru SUZUKI and Dawei Wang

Development of the Mono Ring CVT
Yuichi ITOH and Tomoya SAKAGUCHI

Hybrid Grease NA204F for Automotive Electrical Instruments and Auxiliary Device
Masaki EGAMI, Mitsunari ASAO and Tomoaki GOTO

Ballscrew for Automated Manual Transmission
Koji TATEISHI and Keisuke KAZUNO

Drawn Cup Needle Roller Bearings for Throttle Bodies

Safety Comfort

Hideki AKAMATSU

Compact and Shudderless Constant Velocity Joint (EPTJ)
Tatsuro SUGIYAMA and Yuuichi ASANO

Constant Velocity Steering Joint (CSJ)
Kenta YAMAZAKI

Summary of E Series, Constant Velocity Joint for Propeller Shaft
Tomoshige KOBAYASHI and Teruaki FUJIO

NVH Analysis Using Full Vehicle Multi Body Dynamic Model
Yoshihiko HAYAMA, Takashi NOZAKI, Masafumi NAKAKOUJI, Satoshi FUJIKAWA and Komaki FUKUSHIMA

Development of a High Precision Angle Sensor
Toru TAKAHASHI, Yoshitaka NAGANO and Shoji KAWAHITO

Tapered Roller Hub Bearings for Large Truck
Hiroshi KAWAMURA and Akira FUJIMURAThe Auto-tensioner Market and Technical Trends
Satoshi KITANO, Tadahisa TANAKA and Tomokazu NAKAGAWA

Our Line of New Products

1

2

10

14

20

30

40

48

56

60

66

72

76

80

84

88

92

98

104

110

118

NTN TECHNICAL REVIEW No.73（2005）[ Preface ]Introduction to this Special Feature Issue –NTN's Automotive Products

Kenji OKADAManaging Director

The Kyoto Protocol, adopted in the 3rd session of the Conference of the Parties to the United Nations
Framework Convention on Climate Change (COP3) held in Kyoto in 1997, came into effect in February
2005, and Japan is now facing its obligation to reduce its emissions of greenhouse gases. Recent
automotive technologies are increasingly focused on the challenges of environmental issues. As
evidenced by the rapid growth in the sales of hybrid cars, the auto industry in Japan and around theworld is experiencing dramatic changes in customer preferences. At the 39th Tokyo Motor Show, heldfrom October 21 to November 11, 2005, a variety of novel eco-friendly products and technologies fromvarious car manufacturers around the world were exhibited. As an eco-conscious bearing manufacturer,
NTN exhibited unique products that will, we believe, offer solutions that meet the needs of ourcustomers.
This special issue, NTN Technical Review No. 73 (2005), focuses particularly on three types ofconcerns that automobiles must satisfy – environmental, comfort, and safety. For this issue, we invited
Mr. Iwai Nobuo, Senior Chief Researcher of the Japan Automobile Research Institute FC-EV Center, toprovide an overview of new trends in the automotive industry with a focus on environmental-protection
technologies for automobiles. This issue also summarizes new eco-friendly products and eco-conscious
technologies associated with major NTN products, including constant velocity joints, hub unit bearingsand needle bearings, as well as unit-built and modular products that incorporate electronics
technologies and other unique automotive bearing products.
Our major products, bearings and constant velocity joints, are essentially eco-friendly products that
help reduce energy loss and increase energy efficiency. While constantly seeking quality and safety,
NTN is committed to the development of eco-friendly products that promote greater compactness, lower
torque, higher efficiency and longer service life. In developing these new products, we not only consider
energy savings during use by our customers, but also aim to reduce environmental impacts during
product lifecycles from design, production and sales to use and final disposal. To fulfill these goals, weare making efforts to eliminate the use of substances with harmful environmental impacts in our
products, to introduce highly efficient production lines, to rationalize logistics systems and to achieve
zero waste emissions.
In September 2005, NTN established the General Research & Development Center at our Iwata
Works to further enhance our product development capabilities in high technology fields including next
generation automobiles, environmental protection, energy, medicine and robotics. In the 21st century,
which is being called the century of the environment, NTN is following the motto "For New Technology
Network" in promoting its new R&D projects and contributing to society.

-1-

NTN TECHNICAL REVIEW No.73（2005）[ Contribution ]Technology on Environmental Protection
for The Next Decade

Nobuo IWAI
Senior Chief Researcher FC・EV Center, Japan Automobile Research InstituteExhaust emission purification on both gasoline engine vehicles and diesel engine vehicles could be
solved by 2010 using exhaust catalysts and low sulfur fuels. Global warming gases and energy
diversifications for sustainable mobility will be major issues in the next decade. Use of electric powerdevices will be continued in the major technologies of both hybrid and fuel cell vehicles. Power unit
configuration, traction control four-wheel drive using a traction motor could increase not only energy
savings, but driving safety as well. Advancement of R&D on secure utilities, such as assembly of thetransmission and motor/generator into a single unit is one key technology for environmentally friendly
vehicles.

land would have been further split. To avoid this, theestablished practice was for only the eldest son toinherit the land. A person who violated this traditionand passed his land not only to his eldest son but also
to his younger sons was called "tawake", a foolishperson who splits up his paddy fields. This wise
tradition of our ancestors was intended to preserve
limited assets and resources.
If lands and resources are limitless, however,production expansion is possible and the idea of"tawake" does not apply.
In Europe, the Industrial Revolution offered greater
power, amounts which had not previously been
available from human energy alone. Eventually,
mobility developed by use of transportation means,
typified by automobiles. This revolution in mobility was
triggered by the discovery of an affordable energy
source-petroleum. According to statistical data fromthe Petroleum Association of Japan, global annual
production of this resource reached 413 million
kiloliters in 2004. This great amount of crude oil iscollected from oil fields annually, and a large portion of
it is burned, generating gaseous substances, including
greenhouse gases and particulate matters that are
released into the air. A portion of it is converted into
solids in the form of polymers that are recycled and1. Introduction
Think about the typical fuels and energies used inJapan before the industrialization and westernizationthat took place toward the end of the 19th century.
Average Japanese people used candles andrapeseed oil for illumination, wood and charcoal forcooking and heating, and traveled on foot. Cargo carts
were driven by people, oxen and horses, while boats
and ships were powered by people and wind. In short,renewable energy resources were used in every
aspect of everyday lives in Japan. In those days,
petroleum and coal were available in very limitedareas in small quantities. For nearly three centuries
before industrialization and westernization, Japan
maintained a unique seclusion policy in which thecountry conducted trade with only Holland and China.
During this period, industrial resources and raw
materials were rarely imported into Japan. People
continued to live in a closed land of a limited area on
renewable resources.
In the average farming village, eldest sons inheritedland from their parents. If land had been passed on
not only to eldest sons but also to second and thirdsons, the land would have been split into halves andeven thirds and in the generation of grandsons, the-2-Technology on Environmental Protection for The Next Decade

finally disposed of as refuse or burned and gasified.
Though essential for the present-day economy, as theoccurrence of pollution shows, the development of
petroleum resources also has negative impacts.
Current challenges for automotive environmental
protection technologies are gas emissions reduction
and global warming prevention, while mid and longterm challenges are expected to be related to thesteady acquisition of resources and energies. Thevery purpose of automobiles is transportation, that is,mobility. Maintaining this mobility is necessary even if
the shapes, engines and energy sources of future
automobiles differ from those of current cars. For thispurpose, we need to address the challenges
associated with environmental protection technologies
for automobiles.

during 2007. According to the new 2005 long-term
regulation, the combination of low-sulfur diesel oil andlow-pollution diesel engines will help reduce NOx
emissions to 12% and PM emissions to 3% compared
to conventional diesel-engine cars. Furthermore,according to the 8th Report from the Central
Environmental Council of Japan, NOx emissions will
be reduced to 5% and PM emissions to 1% by 2009.
In other words, virtually all the technologicalchallenges related to automotive gas emissions
control will be solved in the first decade of the 21st
century. Though these gas emissions control
initiatives are limited to the Japanese auto market,
Japanese automakers, having grown into multinational
companies, must now take responsibility forconservation of the global environment. These
Japanese-born environmental protection technologies
will contribute to environmental conservation indeveloping nations where rapid economic growth istaking place. Japanese contributions to internationalsociety in the form of military power and financial
support are often discussed in Japan. However, I
hope that Japan's unique environmental protection
technologies will contribute to the prosperity andwelfare of the world.

2. Gas emissions control
Gas emissions control for gasoline-engine cars has
been nearly solved as the NOx emission level of lowpolluting gasoline-engine cars is now as low as 1.6%of the level of conventional gasoline-engine cars,
thanks to highly advanced exhaust gas catalysts. For
diesel-engine cars that emit not only gaseous
substances including NOx but also particulate matters(PM), emissions will be successfully controlled in thenear future through improved combustion and catalyst
technologies. The petroleum industry in Japandecided to reduce the sulfur content in diesel oil to 50
ppm or lower by the end of 2004 and 10 ppm or lower

3. Recent energy source situation
As can be understood from the statistical data of thePetroleum Association of Japan illustrated in Fig. 1,
the price of petroleum, which was cheaper and100.002005.5

90.00
80.00

Crude oil ($/b)
70.00

LNG (10$/b)

Coal ($/t)

50.00
40.00
30.00
20.00

2004.12Fig.1 CIF basis import prices of crude oil, LNG and coal
(Graph formation using home page data of Petroleum Association of Japan)

-3-

17.3

16.9

16.3

15.9

15.3

14.9

14.3

13.9

13.3

12.9

12.3

11.9

11.3

10.9

10.3

0.00

9.9

10.00

9.3

$

60.00

NTN TECHNICAL REVIEW No.73（2005）

price, the price of West Texas Intermediate (WTI), a
standard US market brand that is produced in Texas,
has always been about $5 per barrel higher. Recently,
the difference between the two prices has been
increasing. For instance, crude oil spot prices as of
October 2004 were $37.54 per barrel for the OPEC
basket and $53.24 per barrel for WTI, a difference of$15.70 per barrel.
This is because WTI is a superior light crude oil that
contains less sulfur content, and because the facilities
that produce light oil from heavy oil lack capacity. As aresult, the supply capacity for light oils, including
gasoline, kerosene and diesel oil, is insufficientcompared to their demand.
According to the estimated amounts of crude oil
deposits and mining years summarized in Fig. 2, thetotal number of mining years left worldwide is 49 and69.3% of proven reserves are in the OPEC nations.
What we should remain aware of is that the mining
years for the USA, China, and Norway and the UK
combined, oil producing nations that also consume
large amounts of petroleum, are 11, 14 and 7 years,
respectivelyﾐsurprisingly small reserves. These mining
years have been obtained by simply dividing theavailable in a greater amount, has been skyrocketingalong with the CIF basis import prices of LNG andcoal. We need to remember the old idea of "tawake,"
which is a theory of finiteness, if this situation
continues for a prolonged period. So far, according to
the statistical data, the CIF based import price of crude
oil has been proportional to that of LNG. However, theCIF based import price of crude oil has not been
synchronous with that of coal. Recently, however, thechange in resource and energy consumption patterns
resulting from dramatic economic growth in developing
nations including China, has resulted in increased
prices for many resources and raw materials. In the
2000-2020 time span, the demand for crude oil in
China seems to have nearly doubled due to China's
dramatic economic growth. Currently, about 90% of
crude oil, the raw material of automotive fuel, comesfrom the Middle East. Likewise, the major sources ofcrude oil for the northeast Asian nations of Taiwan,
South Korea and China are the countries of the Middle
East.
Another uncertain factor about petroleum supply is
the special situation in America. Unlike the OPEC
basket, the crude oil price OPEC uses as a standard■Estimated amounts of crude oil deposits and mining years worldwide (as of the end of 2004)

Unit: million barrels

CanadaAs of the end of 2004, the proved reserves worldwide amounted to approximately 1.28 trillion
barrels with 49 remaining mining years. 69.3% of the proved reserves were accounted for by
the OPEC nations with 57.1% in the Middle Eastern nations.

World total
proved reserves

200
year

1,277,702 (100%)

18
year

4,700
（0.4％）
27,00735,255 39,000（2.1％）（2.8％）（3.1％）

Other Non-OPEC nations

11
year

14
year

7
year

Mexico

41
year

13
year

China

79
year

Norway and UK

Mining
years

USA

87
year

69
year

Saudi Arabia

Kuwait
118
year

Iran

UAE

Venezuela
Libya

Nigeria

96
year

114
year

152
year

Russia37
year

Indonesia

Other OPEC nations

Iraq

Mining years: 49 years

17
year

12
year

12,987 14,600
21,891 18,250 （1.0％）（1.1％）（1.7％）（1.4％）

Provedreserves

77,226（6.0％）

60,000（4.7％）
85,986（6.7％）

97,800 101,500（7.7％）（7.9％） 115,000（8.0％） 125,800（9.8％）OPEC total

Non-OPEC total
178,800（14.0％）

Proved reserves: 885,188 (69.3%)
Mining years: 84 years
Proved reserves: Among the total amount of oil present in a given petroleum
reservoir (original oil-in-place), reserves where production is technically andeconomically feasible are referred to as "recoverable reserves" and are said 261,900（20.5％）
to amount to as much as 20-30% of the "original oil-in-place."
Among projected recoverable reserves, the most reliable ones are referred toas "proved reserves."

Proved reserves: 392,514 (30.7%)
Mining years: 25 years

Mining years: Values obtained by dividing proved reserves at the end of agiven year by the production of the year. For example, "51 mining years"
does not mean a petroleum reserve in the country in question will be
depleted in 51 years because there are possibilities for new well explorations
and excavations, as well as progress in recovery techniques.
Note: The reserve in Canada includes 174.5 billion barrels of oil sand in Alberta.
Source: OGJ magazine (2004 4th quarter issue)

Fig.2 The estimated amount of crude oil deposits and mining years in the world(Home page of Petroleum Association of Japan)-4-Technology on Environmental Protection for The Next Decade

proven reserves by the amounts produced in a given
year. The petroleum supply profile is often said tohave a bell-curve shape. When the peak of the bellcurve is reached and the supply and demandequilibrium breaks, petroleum prices will skyrocket andpetroleum supply will become unstable.

them into common use in order to contribute to energy
conservation and reduction of greenhouse gas
emissions.
If we can reduce energy consumption for production
and use (traveling) by 50%, greenhouse gas
emissions from automobiles will also be reduced by
50% and the mining years for crude oil will be
doubled. The information in Fig. 3, released by theJapanese Ministry of Land, Infrastructure andTransport, summarizes the fuel consumption ofgasoline passenger cars by vehicle weight. As isapparent from this diagram, lighter cars (downsizing)and hybrid cars boost lower fuel consumption.

4. Emission control for greenhousegases and energy conservation
measures for automobiles

Fuel consumption

km/L

Emission of greenhouse gases released from any
country in the world, whether or not the country has
ratified the Kyoto Protocol, can affect not only
neighboring countries but also the global environment.
Currently, only people in a handful of nations, aminority among the living beings on the earth, areenjoying high energy consumption and the benefits ofmobility, and also emitting greenhouse gases in theprocess. Though we tend to think that humaneconomic activities only jeopardize the living
environments of people, we should be aware thatsuch activities also damage the whole of nature.
So that we can continue utilizing convenient
automobiles to maintain mobility while minimizing
greenhouse gas emissions, we must run engines withlow fuel consumption and use sustainable energy
sources to achieve our travel and transportation
purposes. The very task of engineers who specialize
in automobiles is to realize highly efficient engines that
operate on fuels and energies obtained fromresources with stable supplies in order to achieve thisobjective. More specifically, we must realize the costs
and systems that are acceptable to society and bring

40.0
37.5
35.0
32.5
30.0
27.5
25.0
22.5
20.0
17.5
15.0
12.5
10.0
7.5
5.0
2.5
0.0
500

(1) Downsizing
Usually, users prefer heavy, expensive, large luxury
cars. If most hours of car use are for commuting, cars
often carry only one passenger. For example, imagine
a car that weighs 1.5 tons or more traveling with only
one passenger, leaving four or more seats vacant.
Suppose that user purchases a less expensive
lightweight car weighing 1 ton or less, dramatic
reductions in both fuel costs and consumption areachieved simply. For the Academy Awards ceremonyin Hollywood, movie stars arrive in hybrid cars insteadof full-sized luxury cars. This is now regarded as asymbol of status among movie stars. We haveanalyzed the preferences of car users (tendencies
may vary depending on age group or sex) and learned
that certain users are satisfied with smaller cars, or, incertain cases, even motorcycles. For this reason, theauthor believes that we have many possibilities forinvestigation, research and development into
technologies that help realize downsizing.

Conventional engine
Hybrid car
Lean-burn engine
Direct injection engine
Fiscal 2010 target

625

750

875

1000

1125

1250

1375

1500

1625

Vehicle weight

1750

1875

2000

kg

Fig.3 Fuel consumption on gasoline passenger cars
(Home page of Ministry of Land, Infrastructure and Transport)

-5-

2125

2250

2375

2500

NTN TECHNICAL REVIEW No.73（2005）

electric cars is that the distance they can travel on one
charge is limited, a problem that prevents them frombeing used more widely. In Europe, to overcome thisdrawback, a "plug-in hybrid" system is being
considered. These plug-in hybrid cars would function
as externally charged electric cars in urban areas andas internal combustion cars in suburban areas. In
certain situation, they would also operate in hybrid
mode.
There is also the possibility that a novel type ofhybrid car that features a new concept for reducing
greenhouse gas emissions could be developed.
Having two prime movers, that is, an engine and a
motor, hybrid cars can use the driving force of bothwhen a larger output is needed, such as when
accelerating or climbing a slope. As a result, enginesthat are more compact and powerful can be designed
to meet fuel efficiency performance requirements. For
example, the Honda Civic gasoline passenger car has
an engine capacity of 1688 cc and maximum torque of
155 N-m, whereas the Honda Civic Hybrid features anengine capacity of 1339 cc, maximum engine torqueof 119 N-m and maximum system torque of 146 N-m.
In other words, the Honda hybrid boasts a maximum
system torque equivalent to that of the conventional
Civic. In the case of the Toyota Lexus RX300 andHighlander models, the hybrid system consists of the3310 cc engine with a maximum output of 155 kW andfront and rear motors that boost the maximum system
output to as much as 200 kW. Though being SUVs,
these cars feature both the low fuel consumption
associated with popularly priced compact cars and thehigher power of sports cars.
The Toyota Lexus RX300, the Highlander and theEstima Hybrid employ electrically driven 4WD
systems. The front wheels are driven by the engine
and the front motor, while the rear wheels are
powered by the rear motor in response to the current
running state. In the normal state, these cars are
driven by their front wheels. However, during starting,acceleration with full throttle, deceleration and braking,and when slipping is detected on a snow-covered orslippery road surface, the rear wheels also initiate
driving power and the cars switch to 4WD mode.
These cars are good examples of how low fuelconsumption power train technology can be fully
utilized to ensure safe driving.

(2) Hybrid cars
Stationary engines for driving power generators,
etc. can be used as cogeneration systems that offer
exhaust heat and cooling heat for hot water supply. As
a result, such stationary engines boast a total thermal
efficiency of approximately 80%. In contrast, themaximum net thermal efficiency of automotive engines
is usually in the range of 35 to 45% since the waste
heat of these engines cannot be used for hot watersupply. However, in city running modes, typically inthe 10-15 mode, the thermal efficiency drops toapproximately 15 to 17% since partial loads, including
idling with a thermal efficiency of 0%, are often used
and accumulated motion energy is dissipated asthermal energy during braking.
The improved efficiency of hybrid cars is the resultof the following. (1) During deceleration, aregenerative brake uses the accumulated motionenergy to drive the power generator without
dissipating that energy as thermal energy. This
electricity is accumulated and reused for accelerationof the car. (2) The engine stops while the car is notrunning. (3) In low load operations when thermal
efficiency is low, the car runs on electricity that waspreviously generated and stored during medium andhigh load operations when the engine was running at
higher thermal efficiency. If the net thermal efficiency
reaches approximately 30% by these measures, fuelefficiency is doubled.
The description above is the basic concept of hybridcars. To realize this concept, systems that store
energy and convert it into motive power arenecessary. Examples of previous R&D efforts forenergy storage include those for compressed gas andflywheels. Currently, efforts are centered on electricity
storage systems including nickel hydrogen batteries,
lithium ion batteries and ultracapacitors. From theutility viewpoint, the mountability of storage systems isimportant. One of the advantages of hybrid cars overelectric cars is that the size and weight of batteries
mounted in them are less than those in electric cars.
This situation may change if energy storage systems
that are more efficient and useful than electricity
storage systems are successfully developed.
However, electricity storage systems will maintain theirprevalence for the near future.
One of the challenges associated with alleviating
automobile environmental impacts is reducing
greenhouse gas emissions. In regions where a larger
portion of power generation is accounted for by
nuclear and hydraulic power stations, which havelower CO2 emissions, electric cars are more ecofriendly than hybrid cars. However, the drawback of-6-

Technology on Environmental Protection for The Next Decadein order to decrease the price of FCVs. To fulfill thesetargets, government agencies and private
corporations are jointly continuing research anddevelopment activities.
As has been demonstrated in the JHFC project andelsewhere in Japan, FCVs already have performance
sufficient for running in urban streets. Freezing ofwater at cold temperatures is reported to have beensolved by decreasing thermal capacity with a metal
separator and improving the warming capability byusing reaction heat. The operating temperature of thestack is approximately 80˚C. This means smallerdifferences with the outside temperature in the hotsummer climate, so efficient cooling performanceneeds to be provided. If a solid polymer membrane
capable of withstanding higher temperatures isdeveloped, the stack will be able to function normallyat higher operating temperatures, which should makegreater efficiency, improved performance and coolingsolutions easier to attain. The most demanding
challenge for full commercialization of FCVs isachieving a two-digit reduction in price. To realize thisdramatic price reduction, we need to return to the very
basics for research into stack components, such as
solid polymer membranes, catalysts and separators
where dramatic breakthroughs are expected. Anotherchallenge remaining to be solved is increasing therange on one charge of hydrogen. There have been
research efforts and deregulation attempts to increase
the permissible fill pressure of hydrogen containers
from 35 MPa to 70 MPa or greater. However, even if ahydrogen container were charged to 70 MPa, therange of one hydrogen charge would still beinsufficient. A breakthrough is also expected withhydrogen storage technology. For example, highpressure methane hydride stored in high-pressure
containers appears to be promising.
For some years, the possible methods for producing
hydrogen will be onsite production at hydrogen supply
stations and offsite production and shipment frompetrochemical complexes. These will require themodification of petroleum fuels and gases, with cokegas as a byproduct and the electrolysis of water. In
the future, solar heat, a renewable energy, will be
used for steam reforming coal and natural gas toproduce hydrogen without releasing CO2. Japan has
virtually no useable renewable energy sources,
including solar energy. In the future, nations andregions situated in the sunbelt where abundant solar
energy is available, such as the USA, Australia, theMiddle East and Africa, which are also blessed withcoal and natural gas, could become hydrogen supply
sources. Methanol and dimethyl ether seem promising
as carriers for the maritime transportation of hydrogen.

5. In pursuit of sustainable fuelenergies
Another solution to these social needs of energy
conservation and greenhouse gas emissions
reduction that also uses sustainable fuel energies is
the hydrogen fuel cell vehicle (FCV).
In January 2002, the US Department of Energy(DOE) announced the FreedomCAR (Cooperative
Automotive Research) program, which is centered on
the engineering development of hydrogen fuel cells,
as the program to succeed the Partnership for a New
Generation of Vehicles (PNGV). In February 2003,
President Bush announced the Hydrogen Fuel
Initiative that supports the fuel aspect of theFreedomCAR program, and accordingly, ademonstration program has been in progress thatcovers hydrogen production, storage and transport aswell as FCVs. The numerical engineering targets of
the DOE are a durability of 2000 hours for the stack, arange of 250 miles or farther, a hydrogen supply costof $3 per kilogram, and the capability to travel in any
US weather conditions, including low temperaturesand high temperatures with both low and highhumidity, by the end of 2009. The goals for the end of2015 are a stack durability of 5000 hours, a 300-mileor greater range, a hydrogen supply cost of $1.5 per
kilogram, and a fuel cell cost of $30 per kilowatt power
unit. In 2015, the DOE will make a decision onwhether or not to commercialize the FCV, based on
the achievements of these engineering challenges.
Automakers and energy supply companies aremaking joint efforts in fleet tests in various nations.
Examples of such efforts include the California Fuel
Cell Partnership (CaFCP) in the USA, the Clean
Urban Transport for Europe (CUTE) FC bus
demonstration program, and the Ecological City
Transport System (ECTOS) in Europe, and the Japan
Hydrogen & Fuel Cell Demonstration Project (JHFC).
The FCV commercialization scenario of theJapanese Ministry of Economy, Trade and Industry
defines that the period up to the end of 2005 was for
preparation of engineering fundamentals,
demonstration of engineering aspects, establishing
the engineering development strategy, providing legal
frameworks, performing demonstration tests anddeveloping fuel quality standards. In the introductory
phase from 2005 - 2010, the fuel supply system will beestablished, 50,000 FCV units will be introduced to
public organizations and enterprises, and the secondstage FC technology development strategy will be
developed. In the 2010 - 2020 full commercialization
phase, the fuel supply system will be further enhanced
and 5,000,000 FCV units will be sold to ordinary users-7-

NTN TECHNICAL REVIEW No.73（2005）

If this transportation scheme is commercialized,
reformed fuel cells using these carriers as automotive
fuels may be reconsidered.

fuel efficiency improvements in each group.
As this chart shows, in terms of both the averageand top runner fuel efficiency values, the hybrid
vehicles are better than the gasoline vehicles, and thehydrogen fuel cell vehicles are superior to the hybrid
vehicles. Note, however, that this data only shows thedegree of efficiency in the particular period ranging
from the fuel tank to the travel of cars (tank to wheel).
In terms of greenhouse gas emissions, the efficiencyof the whole life of oil, ranging from drilling to fuelproduction and transportation and then operation ofcars (well to wheel) needs to be considered. The
author believes that data on ever-diversifying fuelproduction techniques will be thoroughly reviewed andautomotive engines will be optimized for the fuel types
used.
In a certain sense, fuel cell vehicles are series
hybrid vehicles whose engines and generators are
replaced with fuel cells. Many components, including
secondary batteries and motors, are commonly usedin both fuel cell vehicles and hybrid vehicles. If higher

6. Conclusion
During the next 10 years, hybrid cars will be morecommonly used and fuel cell cars will be increasingly
introduced. Fig. 4 summarizes the comparison of the
10-15 mode fuel consumption of gasoline vehicles(catalog data), hybrid vehicles (catalog data), andhydrogen fuel cell vehicles (actual measurements
obtained from the JHFC project converted to gasoline
equivalents). In each vehicle group, the vehicle
weights have been adjusted to a common value and theaverage 10-15 mode fuel consumption values and top
runner values are indicated. Since there is variation in
the data of actual cars, comparing the average fuelconsumption values and the top runner values (bestvalues available with the present-day technologies) of
all the groups will be useful in order to evaluate theResults of 10-15 mode fuel consumption measurements

Gasoline equivalent fuel consumption [km/L gas.eq.]

FCVs have been verified to boast better fuel efficiency by vehicle weight compared to ICV-HEVs.

40

Gasoline density
Gasoline energy (LHV)
Hydrogen energy (LHV)
Average weight of FCV

41.4: 0.729kg/L: 45.1MJ/kg: 120MJ/kg: 1717kg

38.0
34.5
31.0

30

25.0

23.8

20.7
20
At 55%

10

10.2

11.1

Top
Average runner

At 60%At 65%

Top
Average runnerTop
Average runner0

Catalog data
for ICV group

Catalog data
for HEV group

Actual measurements
for FCV group

Calculated data for FCV models(max. stack efficiency)*

Joint program with NEDO's Solid Molecule Type Fuel Cell System Development Project

As the average vehicle weight, the average for all the vehicles in the FCV group has been taken, and the results for the
ICV and HEV groups have been adjusted to the same average vehicle weight (explained in text).
ICV vehicles: Highlander, X-Trail, CR-V, Zafira, A-Class, Grandis, Wagon R
HEV vehicles: Prius, earlier Prius model, Estima Hybrid, Tino Hybrid, InsightFCV vehicles: FCHV, X-Trail, FCV, FCX, HydroGen3, F-Cell, Mitsubishi FCV, Wagon R-FCV(including vehicles with secondary batteries)*Source: p. 241-264, 2004, "2003 fuel cell vehicles investigation report," Japan Automobile Research InstituteFig.4 Japanese 10-15 mode fuel consumption on gasoline, hybrid and fuel cell vehicles (Home page of JHFC)-8-

Technology on Environmental Protection for The Next Decadeenergy density in energy storage media including
batteries and capacitors is achieved, for example, orother higher functionality is realized, the number ofbatteries mounted on cars can be decreased, allowing
lighter-weight vehicle design.
Furthermore, cost reduction, through improvement
in manufacturing techniques for batteries and othercomponents, is key to realizing greater use of fuel cellvehicles.
One of the technologies that helped create hybrid
vehicles is a unique packaging technology in which
previously independent power units, including anengine, motors, and a transmission are built into
single lightweight assembly that is readily mounted on
a car. The Toyota Hybrid System (THS) in which theelectric CVT is configured using planetary gearing is agood example of overcoming various engineering
challenges.
Currently, an increasing number of models of hybridcars are being introduced into the car market. Someanalysts predict that more than one million hybrid cars
will be produced annually. This technical field naturally
involves bearing technology, which is a strongpoint of
NTN Corporation. The component technology forhybrid cars will be increasingly used for motive powerin electric 4WD vehicles, electrically driven
wheelchairs and construction machinery. NTN's
unique component technologies, including constantvelocity joints, are vital technologies for powertransmission systems on machines including
automobiles. The efforts of the NTN staff to improvetheir technologies to realize more compact,
lightweight, low-loss products support environmental
protection technologies and are reflected in bothconventional and hybrid cars. Examples of such
efforts include their E series high-efficiency constantvelocity joints, PTJ low-vibration constant velocity
joints and mechanical clutch units (MCU).

-9-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Article ]The Reduction of Hazardous Substances

Kiyoshi NAKANISHI**
Masakazu HIRATA**The NTN Group has continued to achieve environmental harmony according to its Environmental
Policy by alleviating environmental impacts arising from its business activities, products and services.
This paper introduces the present and future goals of NTN’s efforts to reduce of hazardous
materials in NTN products. These items would include rolling bearings and constant velocity joints, aswell as many other NTN products.

1. Introduction
Consequently, EU issued a series of directives that
control or ban the use of hazardous substances inthese and other products. These include the ELV (Endof Life Vehicle) directive that totally bans the heavy
metals lead, mercury, cadmium and hexavalentchromium in cars. The RoHS (Restriction of the use ofcertain Hazardous Substances in electrical andelectronics equipment) directive also imposes a total
ban on six hazardous substances, including the heavy
metals above and fire retardants for plastics (PBB's
and PBDE's), in electrical and electronics equipment,
beginning in July, 2006.
In addition, some manufacturers have alreadystarted various self-imposed efforts to ban certain
organic substances that could harm humans.

Believing that coexistence with the global
environment is our most important challenge, NTN
undertakes eco-conscious management aimed at the
reduction of environmental impacts and the realizationof a recycling-oriented society. With these goals, NTN
is striving to reduce CO2 emissions, the prevention of
air, water and soil pollution, the reduction of wastes,
and the promotion of green procurement.
This report describes current and future NTN
efforts to reduce hazardous substances in its
products.

2. Trend in elimination of the use of
specific hazardous substances
In 2001, electronic equipment built in Europe
contained high amounts of cadmium, a hazardous
substance exceeding regulated level standards. This
triggered increased consciousness about green
procurement.

**Environmental Management Dept.
**Research & Development Center Technical Research Dept.

-10-

The Reduction of Hazardous Substances

salt spray test. The tin-plated sheet steel samples
exhibit higher anti-rusting performance compared tothe conventional material.

3. Elimination of controlled substances
3.1 Hexavalent chromium: total ban of use inshields and seal cores

3.2 Lead: total ban of lead-based extreme
pressure agents in grease

Previously, hexavalent chromium had often beenused in NTN product components, including shields
and rubber seal cores on bearings.
To prevent outward leakage of grease and inward
ingress of foreign matter, a shield and a core metalrubber seal are installed on the side faces of sealed
rolling bearings. Conventionally, the metal used inshields and cores has been galvanized sheet steel
and the surfaces of the pieces have been chromatetreated with a material containing hexavalentchromium.
In 2002, NTN introduced a safe tin-plated sheet
steel for shield use with anti-rusting qualities. Twosafe materials for seal core metal to ensure goodadhesion with rubber were also introduced galvanized sheet steel with a special electrolytic
treatment layer and galvanized sheet steel with aphosphate film.
Photo 1 shows examples of tin-plated sheet steel
and conventional chromate-treated (hexavalentchromium) galvanized sheet steel after undergoing a

Lead naphthenate, lead dithiophosphate and leaddithiocarbamate are excellent extreme pressure
agents that have long been used as additives in highload greases.
Constant velocity joints, one of our major products,
are used in high surface pressure and rolling-sliding
conditions. Therefore, greases containing lead-based
extreme pressure additives have often been used onthem to prevent wear and achieve low friction. In
cooperation with a grease manufacturer, we havedeveloped alternative greases to address thisenvironmental challenge1), and started to offer
alternative lead-free greases to our customers in
2002.
Our newly developed greases have unique
combinations of additives that enhance bearing
performance and contain organic molybdenum,
molybdenum bisulfide and similar substances insteadof lead-based extreme pressure additives.

(Conventional) Galvanization + chromating(Improved) Tin plating

5/5 parts

0/5 parts

5/5 parts

5/5 parts

24 hours
later

Rust occurrence

48 hours
later

Rust occurrence<Salt spray test: in accordance with JIS Z 2371> Sample shield plate: 6203 Z, w/ external rust-preventive oil
Test conditions: 5% salt water, spray rate 1-2 ml/h
Test duration: 24 h and 48 h

Photo 1. Shield appearance after (neutral) salt spray test

-11-

NTN TECHNICAL REVIEW No.73（2005）

Performance data for greases developed for theslidable joint TJ series are summarized in Fig. 1 andTable 1. The results for the 3rd order component of
induced cyclic axial load that serves as an index for
NVH (Noise, Vibration, Harshness) performance
affecting automobile riding comfort, as well as fordurability, are better than those of a conventional
grease that contains a lead-based extreme pressure
additive.
In addition, the bearings used in low-speed, highload conditions, such as for steel-making andconstruction machinery, had employed lithium-mineraloil based greases containing lead-based extreme
pressure additives. We have already replaced theseconventional greases with lead-free greases that wedeveloped in cooperation with a grease manufacturer.

4. Efforts to eliminate possible toxic
substances
4.1 Sodium nitrite: ban of use in greases
Sodium nitrite has been used often as a rustpreventive agent because of its excellent performancein this role. It also has high performance as anantiseptic and it is a food additive approved by the US
Food and Drug Administration (FDA). However, it cangenerate nitrosamine, a carcinogenic substance, inthe presence of a secondary amine (Fig. 2).
During the late 1970's, NTN totally banned the useof sodium nitrite-containing coolant as a water-soluble
grinding coolant because grinding machine operators
frequently come into direct contact with such coolants.
Sodium nitrite-containing greases have remained inuse because operators directly touch them lessfrequently, they boast good performance as rustpreventive agents and oxidation inhibitors, and there
have been no reliable alternative greases. Recently,
however, attempts are being made to regulate thistype of grease, especially in Europe. To cope with thistrend, we have already switched most outsourced
greases to alternative safe greases.
In the 1980's, peeling phenomenon accompaniedby the occurrence of a white structure occasionallyappeared on bearings for automotive electricalequipment. This appears to result from hydrogen
brittleness caused by the ingress of hydrogen to thebearings due to the decomposition of the base oils ingreases.

3rd order component of
induced cyclic axial load N (rms)120
Conventional grease
100

Newly developed grease A
Newly developed grease B

80
60
40
20
0

0

100

200

300

400

Torque (Nm)Fig.1 The 3rd order component of induced cyclic axial load of TJ
Table 1 Light load durability of TJ
90

Running time h
125

300

Newly developed No.1
grease A
No.2

○

○

○

○

○

○

Newly developed No.1
grease B
No.2

○

○

○

○

○

○

No.1

△

△

×

No.2

△

△

△

Type of grease

Conventional
greaseT.P

R1

R1
NH + NO2-

R2

N - N =0 + H2O
H＋R2(In oxidizing condition)
Secondary amine

Nitric acid

Nitrosamine

Fig.2 Reaction of secondary amine and nitrite
Target hours(○: No problem △: Minor problem ×: Severe problem)

Table 2 Composition and properties of sodium nitrite free grease

3.3 Cadmium: ban of use in resin colorant

Conventional Newly developed
greasegreaseA yellow or red cadmium-based pigment has beenused often as a colorant for resins. NTN previously
used a cadmium-based pigment in certain resin sliding
material BEAREE products and in the band resin for
expansion-compensating bearings. The cadmiumbased pigment for the former application was replacedwith an alternative material by 1993, and for the latter
application in 2002.

Base oilBase oil viscosity mm2/s
Thickener
Consistency
Life h (at 150°C)
Resistance against brittlenessinduced peeling h
Rust-preventive performance(rust occurrence %)-12-

Ether + PAO
72
Diurea
300
2100

Ether
100
Diurea300
3800＞300（n=5） ＞300（n=11）
1

1

The Reduction of Hazardous Substances

We have verified that sodium nitrite helps create apassive condition on steel material surfaces,
preventing the ingress of hydrogen. Therefore, sodium
nitrite additive is highly effective in preventing this typeof peeling phenomenon. NTN had previously
developed and been using a high-temperature greasethat contained sodium nitrite additives for use inelectrical accessories. To cope with the trend toward
regulation of sodium nitrite, we have developed a new
unique grease. This grease, whose composition andperformance are summarized in Table 2, features
brittleness-induced peeling resistance and rustprevention qualities comparable to those of theprevious grease, with the advantage of longer hightemperature life.

There is no move toward their regulation in Japan,
but in Europe and the USA, attempts are being madeto regulate them.
Rust-preventative oils and greases often containwater-insoluble barium sulfonate. Insoluble in water,
this substance is not yet regulated by the PRTR Law.
However, some researchers suspect that waterinsoluble barium sulfonate can generate water-soluble
salts. Therefore, we are going to introduce rustpreventive oils and newly developed greases that do
not contain barium sulfonate.

5. Conclusion
As stated at the beginning of this report, NTN makes
coexistence with in the global environment a toppriority and implements eco-conscious management
aimed at reducing environmental impacts andrealizing a recycling-oriented society. In this report,
the authors have introduced NTN's proactive efforts to
reduce the use of regulated substances.
Though not described in this report, NTN wishes tocontribute to reducing global environmental impacts
and is fully committed to developing eco-friendly
products that boast smaller sizes, lower torques,
higher efficiency and longer life. Bearings are
essentially eco-products that decrease friction andreduce energy loss. NTN will remain dedicated to thedevelopment of environmentally friendly products andreduction of substances with environmental impacts two areas that are crucial to our eco-conscious
management activities.

4.2 Phthalates: ban of use in boots and rubber
seals for constant velocity joints
Phthalates are often used as plasticizers for plastics
and rubbers. In 2003, certain phthalates were added
to the EU list of regulated substances (EU directive
2003/36/EC) as CMR substances (Substancesclassified as Carcinogens, Mutagens or substances
toxic to Reproduction). More specifically, sales of
dibutyl phthalate (DBP) and diethyl hexyl phthalate(DEHP) and preparations containing these substanceswere banned in the EU region from the end of 2004.*
Certain NTN bearing seals and constant velocity
joint boots contain phthalates as plasticizers. We are
replacing these with phthalate-free components incooperation with associated component
manufacturers with the goal of eliminating phthalates
by the end of 2007.
*Sales of products containing these substances togeneral consumers are permitted.

4.3 Water-soluble barium compounds
Water-soluble barium compounds are listed as
Class 1 designated chemical substances in the
Pollutant Release and Transfer Register (PRTR) Law.

References
1) Shinichi Takabe: NTN Technical Review, 68 (2000), 44

Photos of authors

Kiyoshi NAKANISHI

Masakazu HIRATA

Environmental
Management Dept.

Technical Research Dept.
Research & Development Center-13-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Article ]Application of Topology Optimization and Shape Optimization
for Development of Hub-Bearing Lightening

Haruo NAGATANI*
Tsuyoshi NIWA*

From the perspective of global environmental protection and conservation of resources, thedevelopment of lighter automobile parts for fuel cost savings is a worldwide initiative. In thisabstract, we seek the ultra-lightweight Hub-Bearing, and succeed in its development. In thedevelopment of this product, two optimization methods are used: First is a topology
optimization; the second is shape optimization. We introduce the process of the developmentby explaining the two optimization methods shown above.

the lightest in the world. In the initial stage ofdevelopment, a prototype hub bearing weighing 1.3 kg
failed a durability test, while one weighing 1.4 kg,
illustrated in Fig. 1, succeeded in achieving functionaltargets including for mechanical strength. This result
marked the starting point of our analysis for optimalbearing shape.

1. Introduction
The enactment of the US CAFE (Corporate Average
Fuel Economy) regulations in 1975 triggered
increased demand for lightweight designs for every
automotive component. Lighter components lead tobetter automobile fuel efficiency, which in turncontributes to energy conservation and global
environmental preservation. In manufacturingindustries, various efforts to achieve lightweight
designs are in progress and bearing manufacturers
are committed to decreasing the weights of hub
bearings.1) 2)
NTN has also been working to lighten various
products. By applying shape optimization techniques
to the development of lightweight 3rd generation hub
bearings, we have succeeded in dramatic weight
reductions. This report describes our optimization
techniques as well as examples of lightening effects.

Outer ring
Hub shaft
Inner ring

2. Development goal
The hub bearing being developed is for low weightcars and its target mass is 1.0 kg, which would make it

Fig.1 Hub-Bearing original design*Automotive Sales Headquarters Automotive Engineering Dept.

-14-

Application of Topology Optimization and Shape Optimization for Development of Hub-Bearing Lightening

between a flange and a brake rotor or between an
inner ring and a hub ring are treated. For thisreason, high-precision nonlinear analysis isnecessary to verify the results obtained fromtopology optimization.
(2) Optimal shapes obtained from topology
optimization are complicated 3D forms that may
not be obtainable by machining at reasonable
costs or may not be obtainable at all. Therefore, aprocess to develop a machining-ready shapebased on the results of topology optimization isnecessary. This new basic shape is again
subjected to shape optimization to realize the final
optimal shape.

3. Overview of shape optimization
The basic concept of shape optimization design isto place material in areas that truly need it and thin out
unnecessary material from areas that are notimportant for correct function in order to obtain theminimum shape that satisfies all the necessaryfunctional requirements, such as mechanical strengthand rigidity. Spurred by the lightening needs
mentioned above, recently the demand for the ability
to determine optimal shapes easily has been
mounting. However, currently available software
packages are not powerful enough to fulfill thisdemand.
The currently available analysis techniques can beroughly categorized into (1) topology optimization and(2) shape optimization (these techniques will be
described in detail later). For topology optimization, an
optimal structure for carrying an external load isdetermined, while for shape optimization, the areas tobe altered and the ranges for possible dimensions are
set in advance and optimal values are determinedaccordingly.
Since each type of technique has advantages anddisadvantages, we should utilize the strengths of bothto attain our goals. In our development project,
topology optimization was executed first to develop arough shape (basic shape), and then adjustments
were made as necessary to achieve optimal
dimensions that do not greatly differ from those of thatbasic shape. The reasons for this two-step scheme
are as follows.

*Certain software packages are now capable ofanalysis using gap elements. However, even withthese packages, setup for an object with acomplicated shape is time-consuming, and areas withcomplicated profiles must be set as unalterable. As aresult, these packages are insufficient in terms ofanalytical accuracy.

4. Topology optimization (1st step)In the topology optimization step, an optimal
structure to achieve the functions necessary for thepart in question is obtained. In some cases, dramaticchanges in the structure are possible. We used ahomogenization method 3) as a topology optimization
technique.

Analysis model
The analysis model we used is illustrated in Fig. 2.
The hub ring was integrated with the inner ring, mesh
was plotted and the 3 translational degrees of freedom
for the bolthole nodes on the flange were constrained.
For the external load working on the hub ring, loadconcentration was assumed to occur on the rolling

Necessity of two-step analysis scheme(1) Topology optimization analysis is essentially linear
analysis,* and does not offer sufficiently accurate
results for a boundary nonlinear problem such as
stress analysis for hub bearings where contacts

Hub ring

Flange:design region

Bolt-fastening section:
non-design region

Shaft: non-design region

Inner ring

Wheel spigot joint
outside diameter:
non-design region

Inner side view

Outer side viewFig.2 Topology optimization model-15-

Inside diameter section:
Outside – non-design region
Inside – design region

NTN TECHNICAL REVIEW No.73（2005）

surface and a rolling element load value obtained froma bearing internal force analysis program was utilized.
As shown in Fig. 2, regions where the shape can bealtered (design regions) were set larger, and regions
where the shape cannot be changed (such as thewheel spigot joint outside diameter and bearing
outside diameter) due to interactions with counterpart
members were defined as unalterable regions (nondesign regions).
Furthermore, analysis was executed considering theforging draft, cyclic symmetry conditions of boltholes
and symmetricalness relative to the straight lines
connecting the boltholes and the flange center. The
object function and restricting conditions were as
follows.
Objective function: Distortion energy minimization
Restricting conditions: 22.5%, 27.5% and 30% relative
to the volume (or mass) of initial shape
The analysis was performed with two hub bolt
patterns - 0˚ (cross-pattern) and 45˚ (X-pattern)
relative to the vertical direction. Since the stress was
greater in the 0˚ direction, further analysis was limited
to this direction.

Analysis results

Stress limit value

Maximum stress MPa

The shapes and principal stresses shown in Figs. 3
and 4 were attained. As shown in Fig. 4, with greater
weight reduction, stress rapidly increased. We
adopted a shape that was 27.5% the initial volume
and near the stress limit as the basis for the basic
structure of the hub bearing.

0.1

0.2

0.3

Fig.4 Principal stress of each design

Indicates the initial shape.
(See Fig. 2.)

Wheel spigot joint outside diameter(non-design region)

0.4

Proportion of initial volume

30% the initial volume

27.5% the initial volume

22.5% the initial volume

Fig.3 Result of topology optimization-16-

Application of Topology Optimization and Shape Optimization for Development of Hub-Bearing Lightening

calculate stress values while altering the shapeaccording to the mesh data.
In altering the shape, the shape base vectors* were
set using the morphing software so that alteration of
the mesh could be synchronized with alteration of the
shape in order to maintain the cyclic symmetry of thebolt holes as well as the symmetricalness relative tothe straight lines connecting the bolt holes and theflange center.
The calculation was executed on a 2.8 GHz
Pentium 4 computer and took 2.5 days to complete.
From this result, a response curved surface
approximation model was developed to achieve an
optimal solution.

5. Shape Optimization (2nd Step)
Analysis model
Using the shape obtained from the previously
described topology analysis as a basis, the basic
shape was set as in Fig. 5 and shape optimization
analysis was executed by varying the dimensions at
the seven locations (h, b, y1, x2, y2, x3, y3) shown in
Fig. 6. The fully automatic analysis systemsummarized in Fig. 7 was developed and the analysis
was executed based on the L27 experiment plan
method.
Objective function: Volume (mass) minimization
Restricting conditions: Principle stress at threedesignated nodes (yellow dots in Fig. 6)
below the current level
The software of the shape optimization analysis
system drives the nonlinear analysis solver to*Vectors that defines how each node shifts when
the nodes used as parameters are shifted.

Nonlinear analysis
softwareMorphing software

FEM model

Optimization
softwareDrive
operation
Execution of
nonlinear analysis

Shape base vector setup
Design parameter setup

Write
operation

Read
operation

Objective function:
analysis result

Analysis result

Restricting
conditions

Optimal solution

Fig. 5 Basic design of shape optimization(As installed to the flange of test rig)

Fig. 7 Shape optimization systems

Height: hHeightWidth: b

Fixation at end point of tangential line

Width: by1<Y1
where Y1 is current
value of y1（h,y1,0）X4=const.
h, x2, x3<X4

Y（x2,y2,0）（x3,y3,0）ZX（x4,y4,0）
Fixed

Fig. 6 Parameter of DOE-17-（h,y1,b）（x2,y2,b）（x3,y3,b）

NTN TECHNICAL REVIEW No.73（2005）

L27 best solution

Optimization

Initial value

L27 best solution

Stress point 2

Stress point 1(a) Volume

Optimization

Stress point 3(b) Stress value

Fig. 8 Result of shape optimizationachieved the target mass of 1.0 kg and cleared thetarget values for strength, durability and rigidity. In
addition, the hub ring and outer ring have alsoachieved lightening targets as illustrated in Fig. 10.

Analysis results
In Fig. 8, (a) summarizes volumes, while (b) shows
stress values. "L27 best solution" represents the
combination that performed most effectively in thecalculations performed according to the experimental
method while "Optimization" represents the optimal
solution obtained by calculation in accordance with the
response curved surface approximation. As a result of
this series of calculation operations, many of the
optimal values obtained coincided with the upper orlower limits (limits in terms of avoidance of interaction
with counterpart parts). Consequently, as can beunderstood from Fig. 8(a), there was no apparent
difference between the L27 best solution and the
optimal values obtained from the response curved
surface. We cannot deny the possibility of obtaining
better optimal values, but we believe that we havesuccessfully determined values very near the ideal
optimal values.
Fig. 8(b) summarizes stress values associated withvarying shapes. At every evaluation point, the stressvalue is lower than the current level.
From these findings, we
determined the shape of our
final design.

Fig. 9 Shape of final design

Outer ring

6. Final Design
In the development work, we
also applied the analysis
technique described above tothe outer ring. In addition toshape optimization, we
attempted to achieve lightening
in the finalized bearing
specification through alterationof bearing internal design anddevelopment of new materials
and grease. The newly
developed hub bearing shape,
which is illustrated in Fig. 9, has−100g

Hub ring−240g
Fig. 10 Amount of each component Lightening-18-

Application of Topology Optimization and Shape Optimization for Development of Hub-Bearing Lightening

References

7. Conclusion

1) J. Sakamoto, "Hub unit bearing market andtechnology trends," Monthly Tribology, October 2004
2) K. Kajihara, "Improved hub unit analysis simulation
techniques," Koyo Engineering Journal No. 167(2005)

This report has introduced the shape optimization
analysis executed for lightening hub bearings for
reduced weight cars. With this technique, we havesucceeded in achieving a target mass of 1.0 kg,
attaining dramatic lightening that was previously
considered impossible. Note, however, that this newly
developed hub bearing is not suitable for drum brakestructures in which the hub bearing also functions as abrake seal.* For such applications, another hub
bearing type developed through use of our shape
optimization technique should be used (at a penalty ofadditional 50 g weight).
Our analysis in this report was centered onmechanical strength. However, we also need to set
targets for hub bearing rigidity. Therefore, we will aim
for multi-faceted optimization in which we attempt topromote further lightening while maintaining sufficientbearing rigidity so that we can establish hub bearing
technology to cope with the needs of various car
manufacturers.
*As shown in Fig. 9, the newly developed hub bearing
has four pawls that radially guide the members
installed to the hub ring. If the hub bearing is applied
to a drum brake, the pawls need to be replaced with acircumferential rim to provide sealing function.

Photos of authors

Haruo NAGATANI

Tsuyoshi NIWA

Automotive Sales Headquarters
Automotive Engineering Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

-19-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Paper ]Dynamic Analysis of a High-Load Capacity
Tapered Roller Bearing

Kazuyoshi HARADA*
Tomoya SAKAGUCHI*

It is necessary to predict forces acting on a cage when designing
rolling element bearings. It requires a dynamic simulation that canevaluate interaction forces between the bearing components
including the cage as well as real-time behaviors of thesecomponents.
NTN had already developed a 2-dimensional dynamic analysis
code for cylindrical roller bearings using a commercial, versatile
dynamic analysis software. At this time, NTN has developed a 3dimensional analysis code for tapered roller bearings by extending its dynamic analysis technology.
Additionally, NTN has proposed a tapered roller bearing that accommodates more rollers with a larger outside
diameter cage to increase its load carrying capacity. The cage geometry change may affect the interactions
between the cage and rollers, and the experimental verification of these interactions is generally beyond
accurate measurement. Accordingly, the developed code is implemented to investigate the difference between
the conventional tapered roller bearings and the newly proposed one in terms of the interaction forces between
the cage and rollers and the resulting cage behavior. This report outlines the physical model of this analysis tool,and shows the analytical results where any significant difference is not found from the above viewpoint.
Developers have confirmed that the newly designed high-load capacity tapered roller bearing is quite effective
for use.

NTN has developed a high-load capacity tapered
roller bearing with greater load bearing capacity thatconsists of a cage with a larger outside diameter andan increased number of rollers. The unique geometricshape of this high-load capacity type cage can affectthe interaction between the rollers and the cage, butverification of this problem through experiments isdifficult. Therefore, we investigated the differences in
the cage behavior and the interaction between therollers and the cage by comparing this new type andstandard tapered roller bearings using our newly
developed 3D dynamic analysis tool. As a result, welearned that the new type does not greatly affect thecage behavior or the interaction between the rollers
and the cage.
This report summarizes the 3D dynamic analysis
tool and describes the results of our analysis.

1. Introduction
When designing a cage for a roller bearing,
advanced clarification of the forces acting on the cageis necessary. To analyze the forces acting on thecage, determining the dynamic behavior of the rolling
elements and the cage itself as well as the resultant
forces occurring between the bearing components isnecessary. In short, dynamic analysis is needed.
NTN had previously developed a dynamic analysis
tool for cylindrical roller bearings limited to twodimensional freedom by using the general-purpose
mechanism analysis software ADAMS(R).1) 2) 3) By
enhancing this dynamic analysis technique, we haverecently developed a three-dimensional dynamic
analysis tool for tapered roller bearings.4)*New Product Development R&D Center Mechatronics Research Dept.

-20-

Dynamic Analysis of a High-Load Capacity Tapered Roller BearingΛΛbd

2. Symbolsb: Half the Hertz contact length, mD: Cage outside diameter＿
Diso : Deborah Number =η0eαPu（Gb）/
G: Dimensionless material parameter =α0 E'
E': Equivalent Young's modulus, Pa
Fa: Axial load, N
FEHLr : EHL rolling viscous drag, N
Fpx: Rolling direction component in EHL oil film
pressure, N
Fr: Radial load, Nfc: Cage running frequency, Hz
fT: Friction force at contact area, N
hc: Oil film thickness at center, mk: =1.03αr2/πk': Thermal conductivity of lubricant, W/(mk)＿Lt: Thermal load coefficient =η0βu2/ k'l: Width of roller slice, m
Nsl: Number of roller slices―
P: Mean surface pressure on Hertzian contact, Pa
Pmax : Maximum surface pressure on Hertziancontact, Pa
q: Contact force on sliced piece, NRe: Equivalent radius of curvature, ms: Slip ratio―S: Mean dimensionless shear stress on thewhole contact areaU
: Dimensionless representative velocity＿=η0u（E'R/e）＿u: Mean surface velocity, m/sus: Sliding velocity on contact areaW
: Dimensionless load parameter
Xc: Dimensionless length on EHL contact area=（Diso/Σiso）sinh-1Σisoxc*: Dimensionless x-direction displacement at
cage mass center = x/δPr
yc*: Dimensionless y-direction displacement at
cage mass center = y/δPrzc*: Dimensionless z-direction displacement at
cage mass center = z/δPaα: Pressure coefficient of viscosity, 1/Pa
α0
: Pressure viscosity index of lubricant undernormal pressure, 1/Pa
αr
: Ratio of curvature radius verticallyintersecting the rolling direction to curvatureradius in rolling direction
αp: Pocket angleβ: Temperature increase dependent
coefficient of viscosity, 1/K
δ: Geometric interference amount, m
η0: Viscosity under normal temperature andpressure, Pa･sψ
: [1+2/(3αr)]

Λhd

￣

μbd
μhd
μr
Σiso
σe
δPa
δPr
τ0
φT

: Film thickness parameter = hc /σe: Upper limit film thickness parameter underboundary lubrication: Lower limit film thickness parameter underhydrodynamic lubrication: Friction coefficient under boundary lubrication: Traction coefficient with oil film
: Friction coefficient on contact area
: Dimensionless sheer velocity withisothermal lubricant: Equivalent roughness of contact areas
between two objects, m: Cage axial clearance: Cage pocket radial clearance: Lubricant characteristic stress, Pa: Temperature compensation coefficient of oil film

Subscript charactersb: Roller(s)IR: Equivalent viscosity-solid body modei
: Inner ringPE
: High viscosity-elastic body modeo
: Outer ring
Boldfaced characters represent vectors.

3. Analysis method
The conditions assumed for our analysis are
summarized below.
¡The rollers and cage are provided with six degrees
of freedom.
¡The inner ring is subjected unconditionally totranslational displacement equivalent to rotation at
specific velocities and preset loads (zero degree of
freedom).
¡The outer ring is fixed in space.
¡All the apparent forces, such as centrifugal force,are included.
¡Gravity is considered.
¡Each component is regarded as a rigid body, but
local elastic contact between elements is taken into
account.
¡Interaction force distribution on the roller rolling
contact surface is evaluated with the slice method.
¡For the friction force between the rollers and the
raceway, the friction component resulting from theoil film and the metal contact is considered. Also, inthe elastohydrodynamic lubrication (EHL) condition,
the rolling viscous resistance5) is considered (Fig.
1).
¡The squeeze effect of the EHL film (speeddependent term) is not considered.
¡All the interaction force between the roller large end
face and the inner ring large rib face is assumed to-21-

NTN TECHNICAL REVIEW No.73（2005）

Contact and Tangential (Traction or Friction) forces
EHL Rolling ResistanceContact andFriction forces

Contact and Tangential (Traction or Friction) forces
EHL Rolling ResistanceContact andFriction forces

Contact andFriction forces
Contact andFriction forces

Contact and Tangential(Traction or Friction) forces

Contact andFriction forces
Contact and Tangential (Traction or Friction) forces
EHL Rolling ResistanceContact and Tangential (Traction or Friction) forces
EHL Rolling ResistanceFig. 1 Considered interaction forces in this dynamic analysis

be applied to the maximum neighboring point. The
friction coefficient is handled in the same manner asthat of the raceway surface. However, because of
the sliding contact, the EHL rolling viscous
resistance is not considered (Fig. 1).
¡For the friction force between the rollers and thecage, boundary lubrication alone is assumed (Fig.
1). In the case of contact with the roller end face, all
the contact force and friction force of the maximum
interaction point is applied.

μbd
μr =

μbd−μhd（Λ−Λhd）6＋μhd if Λbd≦Λ＜Λhd6（Λbd−Λhd）
if Λhd≦Λ

μhd………………………………………………
（2）u
ft=μr q・ s …………………………………（3）us
For this purpose, Λbd=0.01 and Λhd=1.5 wereused. The above-mentioned film thickness
parameter is essentially the center oil film thickness.
The center oil film thickness was determined byassuming that the contact area was in thehydrodynamic lubrication mode (either the highviscosity-elastic body mode (PE mode) or theequivalent viscosity-rigid body mode (IR mode), as
necessary) as described in expression (4). For the
PE mode, Pan's formula (5)6) was used while for the
IR mode, Martin's formula (6)7) was used.
Furthermore, considering the inlet temperature,
the temperature compensation coefficient φT 8)
defined by expression (8) was used. Usually, the IR
mode is necessary when the rollers are in contact
with the outer ring in the non-load region.
Additionally, through comparison of the oil film
thicknesses of the two fluid lubrication modes, therelevant mode was determined, and this mode wasused to judge whether the EHL rolling viscous
resistance, described later, was necessary.

3.1 Dynamic interaction forces between theroller contact surface and the raceway
surface
The dynamic interaction force model for theinteraction between the roller contact surface and the
raceway surface is based on the following assumptions.
1) Only when an elastic deformation amount(geometric interaction amount δ) occurs against
the relative position of the rollers and the bearing
ring that is governed by the time t, a contactinduced normal force, friction force and rollingviscous resistance occur between the roller rolling
contact surface and the raceway surface.
2) The slice method is applied to the roller rolling
contact surface, thereby the distribution of theinteraction force is considered. If the interaction
vector δ on each slice is given, then Palmgren's
simple formula is employed and the expression (1)is used to determine the contact force vector q foreach slice.
q=0.356E'Nsl-1/9 l 8/9δ10/9

if Λ＜Λbdδ………………（1）δhc=3) The friction coefficient is calculated for eachlubrication mode in the manner shown byexpression (2), based on the film thickness
parameter.3) The friction force vector fT acting oneach roller slice is determined with expression (3)using the roller-dependent sliding velocity vector us.

φT・hc, PE

if hc, PE＞hc, IR

φT・hc, IRotherwise

…………
（4）

（5）
hc, PE=2.922 W-0.166 U0.692 G0.47 Re …………
hc, IR=4.9 UW-1 Re-22-………………………（6）Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing

The mixed lubrication region (μbd±Λ＜μhd) was
determined by smoothly interpolating the friction
coefficients μhd andμbd of the above-mentioned
fluid lubrication and boundary lubrication,
respectively, as shown by expression (2). Fig. 3
shows two cases of variation in friction coefficients,
withμhd of 0.001 and 0.06, determined throughexpression (2).
4) To determine the EHL rolling viscous resistance,
Zhou's regressive formula (13) was used, but thisoperation was not performed when the applicablefluid lubrication mode was the IR mode. The force of
EHL rolling viscous resistance was assumed to be＿
opposite that of the mean surface velocity vector u
Also, with two rotating objects, it is necessary toconsider the force Fpx (expression (14)) that resultsfrom the pressure component in the rolling direction
on the EHL oil film.5) Note, however, that the sign for
the rollers and inner ring must be a plus sign, andthat for the outer ring must be a minus sign. Being
compensated for by the component force ofpressure in the normal direction, this force does notaffect the moment.

Provided that,

Re-1=

1
1
for Roller / Inner race
＋
Rb
Ri
1
1
for Roller / Outer race
−
Rb
Ro

φT =

1−13.2（Pmax /E'）LT0.42
1＋0.213（1＋2.23s0.83）LT0.64…（7）………（8）For the friction coefficient μbd under the boundary
lubrication condition (Λ<Λbd) shown by theexpression (2), the function9) of Kragelskii's solid
contact friction coefficient was modified and used asshown in Fig. 2. The modification was such that thevariable was changed from a sliding velocity to asliding parameter and the friction coefficient was set
to 0 when sliding was also 0.
（9）μbd =（−0.1＋22.28s）exp（−181.46s）＋0.1…
The friction coefficientμhd for the fluid lubrication
condition (Λ<Λhd) was calculated with Muraki's
simple theoretical formula10) given below. To reduce
the amount of numerical calculations, thetemperature was assumed to be constant. In
addition, it was assumed that this traction model
could also be effective even in the IR mode.

FEHLr=

29.2Re（l GU）0.648 W0.246 u・uα0

if hc, PE＞hc, IR
otherwise0………………………………………………
（13）

Xc≧2：S=Σiso/Diso ………………………（10）Fpxb,r=±Xc＜2：S=sinh-1Σiso｛1−（Diso/4Σiso）sinh-1Σiso｝

u…………………………（14）
u

2Re
FEHLr
Rb,r

As such, the contact force, friction force and EHL
rolling viscous resistance acting on each roller slice
dynamic contact surface were calculated. In
addition, the moment on each slice was calculated.
The total of forces and moments working on all theslices act on the roller.

………………………（11）μhd=τ0 S/P ………………………………（12）Here, the dimensionless length of the elastic area
is represented as Xc = ( Diso/Σiso) sinh-1Σiso.

0.15

0.15

0.12

Friction coefficient μr

Friction coefficient

−φT

0.09
0.06
0.03
0
0

0.01

0.02

0.03

0.04

0.12

0.09

μhd =0.06

0.06

μhd =0.001

0.03

0.05
0
0.001

Slip ratio

Fig. 2 Friction coefficient under boundary lubrication

0.01

0.1

1

10

Oil film parameter Λ

Fig. 3 Relationship between friction coefficient and oilfilm parameter-23-

NTN TECHNICAL REVIEW No.73（2005）
Large endside

3.2 Interaction forces between the roller large
end face and the inner ring large rib face
Generally, the contact area between the roller large
end face and the inner ring large rib face has 30-40%
slip. For this reason, the rolling viscous resistance isapproximately one tenth of that of the traction force orfriction force and can therefore be ignored.
1) The roller large end face of the bearing is roundshaped and the inner ring large rib face is coneshaped. As a result, elliptical contact occurs.
Because the length of elliptical contact is smaller
than the contact length between the roller and the
raceway surface, it can be assumed that all theinteraction forces act on the geometrical maximum
interaction point.
2) From the geometrical interaction amount δ, thecontact force can be calculated with Hertz's point
contact formula. The friction coefficient for the roller
large end face and inner ring large rib face can becalculated in a manner identical to that for the
raceway surface. Note, however, for the calculationof oil film thickness, Brewe's11) formula (16) is usedin the IR mode, and Chittenden's12) formula (17) isused in the PE mode, as described below.
hc=

φT・hc, PE

if hc, PE＞hc, IR

φT・hc, IRotherwisehc, IR=128αr 0.131tan-1αr＋1.683
2

Small end side

Fig. 4 Geometrical shape of cage pocket

Because the roughness of the pocket surface is great,
the friction force when the roller rolling contact surface
is in contact with the cage pocket surface wascalculated assuming that the lubrication mode wasboundary lubrication. In addition, because the width of
the bar of the pocket in contact with the roller rolling
contact surface is finite, a case where the roller was incontact with the edge of the bar was considered.

3.5 Interaction forces between the roller end
face and the cage pocket surface
The interaction forces between the roller large end
face and the cage pocket surface and between theroller small end face and the cage pocket surface arein sphere-to-plane and plane-to-plane contact modes,
respectively. Therefore, we subjected all the
representative points where contact could occur to aseries of calculations, and took the sum of the contact
and friction forces at each point as the interaction
force at that point. Because the roughness of thepocket surface is great, the friction force when theroller end face was in contact with the cage pocket
surface was calculated assuming that the lubrication
mode was boundary lubrication.

……………（15）22ψDetail of pocket

U R
e ……
（16）
W

（17）
hc, PE=4.31U 0.68 G 0.49 W -0.073 1−exp（−1.23k2/3） Re…

3.3 Interaction forces between the roller small
end face and the inner ring small rib face
Generally, contact between the roller small end face
and the inner ring small rib face rarely occurs.
However, if gravity and external vibration act on thisarea in the non-load region, the roller small end facecan contact the inner ring small rib face on rareoccasions. Again, assuming that the entire interaction
force is applied to the maximum contact point, we
calculated the interaction force using Hertz's point
contact formula. Because the mode of contact in thisarea was considered to be edge contact, we used afriction coefficient under constant boundary lubrication
for the friction force calculation.

4. Analysis model
The analysis model and coordinate systems used
are shown in Fig. 5. Every coordinate system used
was a right-hand coordinate system. Two types ofbearings "standard specification and high-load
capacity" were subjected to dynamic analysis todetermine the differences in the cage behavior andinteraction force between them. The high-load
capacity type has inner and outer rings and rollers
whose diameters were the same as those on thestandard specification type, but its newly developed
cage has an outside diameter greater than that of thestandard specification bearing. The bearing data andoperating conditions for these bearing types are
summarized in Table 1, and the area that includes thedimensions in Table 1 is illustrated in Fig. 6.
The radial load was applied such that the inner ring
was displaced in the +y direction and the topmost
roller received the maximum load.

3.4 Interaction forces between the dynamic roller
contact surface and the cage pocket surface
The interaction force between the dynamic roller
contact surface and the cage pocket surface (shownwith diagonal lines in Fig. 4) is similar to theinteraction force between a roller and a bearing ring,
and, therefore, was evaluated with the slice technique.
-24-

Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing

5. Analysis results
In further analysis, to eliminate the influence of theinitial conditions the data acquired during the 0.5 s
time-span after the start of calculation (forapproximately 40 revolutions of the inner ring) was
deleted, and the data obtained during the next 0.2 s(for approximately 17 revolutions of the inner ring) was
utilized for evaluation.

5.1 Translational displacement on the radialplane at the cage mass centerOn the cage used for the analysis, the pocket radial
clearance is larger than the pocket circumferential
clearance as illustrated in Fig. 6(b), and the maximum
translational displacement of the cage is the radial
clearance on the pocket 4). The trajectories of the cage
mass centers on their radial planes with the standard
bearing and the newly developed bearing are plottedin Fig. 7. The red solid lines in the charts represent
the radial plane positions of the cage mass centersthat were made dimensionless with the radial pocket
clearances. The blue dotted lines indicate circles thattook the radial pocket clearances of the cages as aradius. The areas of trajectories of cage mass centers
are smaller than the radial pocket clearances, andmatch the characteristics of cage behavior undercomplex load conditions described in previous
research4, 13). The displacement of the newly
developed cage is slightly smaller than that of thestandard cage.
The pattern of interaction force between the roller
rolling contact surface and the cage is illustrated inFig. 8. Fig. 8(a) shows the cage/roller interaction force
and the interaction position in the case where thecage mass center is displaced to the position shown in
Fig. 8(b). The red vector in this chart is aninterference force which the cage exerts onto therollers and originates from the center of gravity ofeach roller. As can be understood from this chart,

Fig. 5 Analyzed bearing schematic and its coordinate system
Table 1 Test bearing and operating conditions
Standard specification
High-load
bearing(inside dia.×outside dia.×width, mm) （φ40×φ76.2×17.5） capacity bearing*

Bearing

Cage type
Number of rollers
Basic rated dynamic load, kN
Pocket angle αp, deg
Cage outside dia φD，mm
Lubricant temperature ˚CLubricant dynamic viscosity, mm2/s {cSt}
Inner ring running speed, rpm
Load, kN

Standard cage
21
46.5
45
61.39

Newly developed cage
23
50.0
56
62.22

100
2.52 @ 100˚C
5000
Fr = 5，Fa = 2.5*The inner ring, outer ring and rollers are identical to those used on the

φD

Pocket radial
clearance

standard specification bearing.

Pocketcircumferential
clearanceαp(a) Assembled bearing(b) Cross-sectional viewof cage pocket

Fig. 6 Cage geometry

Radial Pocket Clearance
1.0

Displacement yc＊

Displacement yc＊

1.0

0.5

0.0

-0.5

-1.0
-1.0

-0.5

0.0

0.5

0.5

0.0

-0.5

-1.0
-1.0

1.0

＊

-0.5

0.0

0.5

Displacement xc＊(b) Newly developed cage

Displacement xc(a) Standard cage

Fig. 7 Numerical results of cage mass center behavior in radial plane

-25-

1.0

NTN TECHNICAL REVIEW No.73（2005）

Displacement yc＊

1.0

0.5

0.0

-0.5

y

Position of cage mass center
-1.0
-1.0

x

-0.5

0.0

0.5

1.0

Displacement xc＊(b) Position of cage mass center(a) Roller/cage interference position

Fig. 8 Relationship between the direction of cage displacement and cage/roller interaction position (standard cage)

the newly developed cage. The charts in Fig. 9 wereobtained by plotting the absolute values of theinteraction forces between the dynamic roller contact
surfaces and all the cage bars. The interaction
between the dynamic roller contact surface and thecage appears to be non-cyclic. The mean andmaximum values of the interaction forces are
summarized in Fig. 10. The mean value with the
newly developed cage is 14% smaller than that withthe standard cage. We believe this is because thenormal load from the raceway surface on each roller

Cage/Roller Force, N

100.0

0.0
0.5

100.0

0.0
0.6

0.7

0.5

0.6

Time, sTime, s(a) Standard cage(b) Newly developed cage

Fig. 9 Cage/roller (rolling surface) interaction forces
50

0.3

45

35

0.2

30
25

0.15

20
0.1

15
10

Mean force, N

0.25

40

Maximum force, N

Cage/Roller Force, Nunder this operating condition, multiple rollers that aresquare to the displacement direction of the cage areinteracting with the cage.
In this situation, the direction of displacement within
the pockets of the rollers in contact with the cage isthe circumferential direction shown in Fig. 6(b). As aresult, the displacement of the cage is smaller thanthe pocket radial clearance.
A time-dependent record of the interaction force
between the dynamic roller contact surface and thecage is given in Fig. 9 for both the standard cage andMaximum force
Mean force

0.05

5
0

Current
bearing

Developed
bearing0

Fig. 10 Mean value and the maximum value of cage/roller (rolling surface) interaction forces

-26-

0.7

Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing

has decreased due to the increased number of rollers.
The greater number of rollers has caused the traction
force from the raceway surface that drives the rollers
to decrease and resulted in decreasing the meaninteraction force between the dynamic roller contact
surface and the cage.
The driving force from the cage stems from theinteraction force on the rollers, and the timeintegration value of this interaction force governs the
behavior of the cage. For this reason, the difference inthe amplitude of the mass center trajectories of thecages, as shown in Fig. 7, seems to result from thedifference in the mean interaction forces.
The difference in the maximum interaction forces ofthe standard cage and the newly developed cage was
2%. The maximum interaction force can be affectedby accidental interaction. However, when all the dataobtained from approximately 57 revolutions of theinner ring was evaluated, the maximum interaction
force with the newly developed cage did not exceedthat of the standard cage by more than 2%. Thus, we
can judge that the maximum interaction force isroughly the same for both cage types.

5.2 Axial displacement of the cage mass center
Next, the time-dependent trend of the axial
displacement of the mass centers of both cages is
illustrated in Fig. 11. As shown in Fig. 12, the verticalaxis took the cage mass center as the zero point when
the roller was in contact with the raceway surface andthe large rib face of the inner ring and the cage werein contact with the roller large end face. The verticalaxis was also made dimensionless with the axial
clearance of the cage pocket. As can be seen in Fig.
11, the axial displacement of the cage wasapproximately 1/2 the maximum axial clearance of the
cage. In Fig. 13, cage/roller interaction force patterns
in the cage pocket are illustrated. The red arrow in the
diagram shows the interaction force from the cageacting on the roller. As a result, a reaction force in
response to the illustrated interaction force acts on the
cage. Due to the geometrical shape of the cage
pocket, if the dynamic roller contact surface interfereswith the cage pocket (Fig. 13(a)), the cage shifts inthe -z direction due to the z-direction component of
the above-mentioned interaction force. On the other
hand, when the roller large end face interferes with thecage pocket (Fig. 13(b)), the cage shifts in the +z
direction.

1.0

Displacement, zc＊

Displacement, zc＊

1.0
0.8
0.6
0.4
0.2
0
0.5

0.55

0.6

0.65

0.7

0.8
0.6
0.4
0.2
0
0.50

0.55

0.6

0.65

Time, sTime, s(a) Standard cage(b) Newly developed cage

Fig. 11 Axial behavior of cage mass centerzc
Contact areas

Fig. 12 Datum point for axial displacement of cage mass center

-27-

0.7

Cage
revolution
direction

Small end face

X

Large end faceZ

Small end face

Large end face

NTN TECHNICAL REVIEW No.73（2005）

X

Interaction force
from cage acting
to rollerZ

Cage
revolution
direction

Interaction force fromcage acting to roller(a) Contact between roller rolling surface and cage pocket(b) Contact between roller large end and cage pocket

Fig. 13 Graphic example of cage/roller interaction forces (standard cage)

The roller large end face/cage pocket interaction
force is summarized in Fig. 14. This chart was
obtained by plotting the absolute values of theinteraction forces between the large end faces of therollers and the cage pockets. On the standard bearing,interaction occurs at about the cage running frequency
fc while on the newly developed bearing, themagnitude of the interaction force is approximately 1/2that of the standard bearing and this means theinteraction occurs more frequently on the newly
developed cage. In terms of axial displacement, thedifference between the standard cage and the newly
developed cage in Fig. 11 results from the difference
between interaction patterns given in Fig. 14. Under
our analysis conditions, no collision between the roller
small end face and the cage pocket occurred.
As described above, the behavior of the newly
developed cage is the same as that of the standard
cage, while the roller/cage interaction force with the
newly developed cage is smaller than that with thestandard cage. The effect of the cage outside
diameter and pocket angle on the roller/cageinteraction force is small within the scope of thisanalysis.

6. Conclusion
Using a 3D dynamic analysis tool optimized fortapered roller bearings, we have compared a standard
tapered roller bearing and a high-load capacity
tapered roller bearing that has an increased number ofrollers. As a result, we have found that the behavior of
the cage with the high-load capacity bearing is very
similar to that of the standard bearing and that theroller/cage interaction force with the high-load capacity
bearing is equivalent to or smaller than that of thestandard bearing. We also found that the effect of thegeometrical shape of the cage unique to the high-load
capacity bearing onto the roller/cage interaction force
is small within the scope of our analysis.
It is difficult to experimentally compare the levels ofroller/cage interaction force. Therefore, our analysis
technique, as an alternative to an experiment-based
technique, is a useful means for verifying the functionsof high-load capacity bearings.

10.0

Cage/Roller-end force, N

Cage/Roller-end force, N

10.0
7.5

5.0
2.5
0
0.5

0.55

0.6

0.65

0.7

7.5

5.0
2.5
0
0.5

0.55

Time, s0.6

0.65

Time, s(a) Standard cage(b) Newly developed cage

Fig. 14 Cage/roller large end interaction forces

-28-

0.7

Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing

References
1）MSC.Software，HP Address：
http://www.mscsoftware.co.jp/（2005.05.23）
2) Tomoya Sakaguchi, Kaoru Ueno, Takuji Kobayashi:
Analysis of cage behavior on cylindrical roller
bearings, Japanese Society of Tribologists, Tribology
Conference Preprints (Sendai 2002-10) 415.
3) Tomoya Sakaguchi, Kaoru Ueno: Analysis of cage
behavior on cylindrical roller bearings, NTN
Technical Review, No. 71 (2003) 8-17.
4) Tomoya Sakaguchi, Kazuyoshi Harada: Analysis ofcage behavior on tapered roller bearings (2nd report,
calculation result), Japanese Society of Tribologists,
Tribology Conference Preprints (Tottori 2004-11)503.
5）Zhou, R. S., Hoeprich, M. R.：Torque of Tapered
Roller Bearings, Trans. ASME, J. Trib., 113, 7（1991） 590-597.
6）Pan, P., Hamrock, B.J.: Simple Formulae for
Performance Parameters Used in
Elastohydrodynamically Line Contacts, Trans.
ASME, J. Trib., 111, 2（1989） 246-251.
7）Martin, H. M., :Lubrication of Gear Teeth,
Engineering, London, 102（1916）119-121.

8）Gupta, P. K. et al., : Visco-Elastic Effects in Mil-L7808 Type Lubricant, Part I; Analytical Formulation,
STLE Tribol. Trans., 34, 4（1991） 608-617.
9）Kragelskii, I. V., : Friction and Wear, Butterworths,
London（1965）178-184.
10) Masayoshi Muraki, Yoshitsugu Kimura: Research into
traction characteristics of lubricant (2nd report),
Junkatsu, 28, 10 (1983), 730-760.
11）Brewe, D. E., Hamrock, B. J., Taylor, C. M., : Effectsof Geometry on Hydrodynamic Film Thickness,
ASME J. Lubr. Technol., 101,2（1979） 231-239.
12）Chittenden, R. J., Dowson, D., Dunn, J. F., Taylor,
C. M., : A theoretical analysis of the isothermal
elastohydrodynamic lubrication of concentrated
contacts I. Direction of lubricant entrainment
coincident with the major axis of the Hertzian contact
ellipse, Proc. Roy. Soc., London, A397（1985）245269.
13) Kazuyoshi Harada, Tomoya Sakaguchi: Analysis ofcage behavior on tapered roller bearings (1st report,
behavior measurement), Japanese Society of
Tribologists, Tribology Conference Preprints (Tottori
2004-11) 501.

Photos of authors

Kazuyoshi HARADA

Tomoya SAKAGUCHINew Product Development
R&D Center
Mechatronics Research Dept.

New Product Development
R&D Center
Mechatronics Research Dept.

-29-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]High Capacity Tapered Roller Bearings
・High Rigidity Tapered Roller Bearings - Super Low Torque・

Takashi TSUJIMOTO*
Jiro MOCHIZUKI*

Tapered roller bearing have greater capacity for carrying not only pure radial or axial loads but also
combined loads, and feature greater bearing rigidity. Therefore, they are found in numerous
applications in various industries such as the automotive industry.
Recent advancements in transmissions for low fuel consumption have resulted in lower oil
viscosities and reduction in transmission size. Therefore, it is necessary to reduce the size of the
bearings, which can result in bearing life and rigidity problems.
High capacity tapered roller bearings were developed in order to suppress the life reduction effectassociated with reduction in bearing size. This is accomplished by increasing the number of rollers(similar to a full compliment type bearing) by using a special cage. By doing this, the dynamic load
rating can be increased by up to 10%, and the static load rating can be increased by up to 15%. As
a result, bearing life under severe lubrication conditions can be improved in addition to increasing thebearings rigidity.
Super low torque bearings that maintain bearing life and rigidity can be designed by combining
high capacity tapered roller bearings and FA tapered roller bearing features.

1. Introduction
In order to reduce automobile fuel consumption,
efforts are in progress to introduce low-viscosity oils
and compact, lightweight designs for automotivetransmissions and differential gears. The sizes ofbearings used in transmissions and differential gears
have been getting much smaller, and, as a result, theneed to ensure bearing life and rigidity has been
increasing. To address this challenge, we havedeveloped a unique high-load capacity, compact
tapered roller bearing that has a usage life just as longas traditional bearings.

This bearing type incorporates a special cage andhas an increased number of rollers, nearly the same
number as a full complement roller bearing, toincrease the load carrying capacity. As a result, thecontact pressure is decreased, higher bearing rigidity
is achieved and bearing life is extended even under
severe lubrication conditions, including contaminatedlubricant. In this report, the authors introduce thestructure and features of this bearing and the resultsof evaluation tests, as well as examples ofapplications of this low-torque, high-rigidity design.

*Automotive Sales Headquaters Automotive Engineering Dept.

-30-

High Capacity Tapered Roller Bearings

2. Structure of high-load capacity
tapered roller bearing

standard bearing. At the same time, a narrower rollerto-roller clearance was adopted so that the bearing
cage can incorporate almost as many rollers as thenumber used in full complement bearings.

Employing a smaller clearance between the cage
and the outer ring to increase the cage PCD allows
the cage bar width to remain the same as that of aStandard bearing

High-load capacity bearing

Clearance between
outer ring and cage

Clearance between
outer ring and cage(smaller)Cage bar width

Clearance
between rollers

Cage bar width(same as conventional
bearings)

Clearance
between rollers(smaller)

Fig. 1 Structure of high capacity tapered roller bearing

Standard bearing

High-load capacity bearing

24 rollers
Dynamic rated load Cr= 42.0 kN
Static rated load Cor= 52.0 kN

27 rollers (increase of 3 rollers)
Dynamic rated load Cr= 45.5 kN (increase of 8%)
Static rated load Cor= 58.0 kN (increase of 11%)

Fig. 2 Example of bearing design (size : φ45×φ81×16)

lubricating conditions
The greater number of rollers contributes toreducing the maximum contact surface pressure,increasing the oil film thickness, and stress relief
in metal-to-metal contact situations. Thus,
surface-initiated flaking, which results from metalto-metal contact in lubricating conditions when an
oil film is not readily formed, is inhibited resulting
in the extension of practical bearing life.
¡Improved practical bearing life undercontaminated lubricating conditions
The increased number of rollers contributes toreducing the maximum contact surface pressure,thereby decreasing the size of dents caused bytrapped foreign matter are and the stressoccurring on the rims around dents is reduced. As
a result, the bearing life under contaminatedconditions is also extended.

3. Features of high-load capacity
tapered roller bearing
The increased number of rollers gives our high-load
capacity tapered roller bearing the following improved
functions:1) Increased rated load
¡Up to 10% increase in dynamic rated load
(maximum improvement of 37% in terms of
calculated life)¡Up to 15% increase in static rated load (maximum
improvement of 15% in terms of safety)

2) Higher rigidity¡Up to 10% increase in bearing rigidity (up to 9%
decrease in elastic displacement)

3) Longer bearing life
¡Improved practical bearing life under clean-31-

NTN TECHNICAL REVIEW No.73（2005）

Λ*: Oil film parameter (oil film thickness/combinedrolling contact surface roughness)

4. Performance of high-load capacity
tapered roller bearing

2) Results of bearing life test for use withcontaminated lubricant

Table 1 Test bearing
Standard bearing
Bearing size

The results of the life test under use withcontaminated lubricant are summarized in Fig. 4.
Our high-load capacity tapered roller bearing wasless prone to develop dent-initiated flaking in use withcontaminated lubricant and boasts a bearing life
approximately 3 times longer than standard tapered
roller bearings under the same conditions.

High-load capacity bearing

φ45×φ81×16 mm

Dynamic rated load Cr =42.0 kN Dynamic rated load Cr =45.5 kN
Rated load
Static rated load Cor=52.0 kN Static rated load Cor=58.0 kN
24 rollers
27 rollers
Number of rollers

3) Results of bearing life test in use with lowviscosity lubricant

1) Results of a bearing life test under a clean
lubricating condition of Λ*0.2

Generally, the viscosity grade of the oils currently
used on automotive MT and differentials is in therange of VG75 to VG90, while the viscosity level ofoils on automotive CVT and AT is at the VG32 level.
The lowest viscosity grade oils currently available on
the market are VG20 level oils. Because transmission
oils with much lower viscosity will be increasingly
used, we executed a bearing life test using viscosity
grade VG10 oil under both clean oil lubrication andcontaminated oil lubrication conditions.

Accumulated failure probability, %Fig. 3 illustrates the results of a bearing life test
under Λ*0.2 clean oil lubricating conditions in which
oil film formation was extremely poor.
Our high-load capacity tapered roller bearing wasless prone to developing surface-initiated flaking,which results from metal-to-metal contact in poor oilfilm formation conditions, and boasted a bearing life
approximately 15 times longer than standard tapered
roller bearings.

Load conditions:radial load Fr= 19.1 kN
axial load: Fa= 0 kN
(axial clearance 0.20-0.25 mm)
Speed: 2500 min-1
Lubricant: Jomo High Speed Fluid VG1.5
Lubrication method: oil bath
Oil film parameter: Λ = 0.2
Calculated life:
standard bearing L10= 92.2 h
High-load capacity bearing L10= 120.4 hStandard bearing
High-load capacity
bearing

95
90
80
70
60
50
40
30

slope e20

L (50)

1.546
4.231

22.4
152.9

10

L (10)

5

6.6
98.0
3

5

7 1
2
×100

3

5 7 1
2
×101

3

5 7 1
2 3
×102

5 7 1
×103

Life, hLoad conditions:radial load: Fr= 19.1 kN
axial load: Fa= 0 kN
(axial clearance 0.20-0.25 mm)
Speed: 2500 min-1
Lubricant: Turbine Oil VG56
Contaminants: steel beads + gas-atomization 0.5 g/LContaminant grain size
50μm or smaller, 90wt%
100-180μm, 10wt%
Lubrication method: oil bath
Oil film parameter: Λ = 2.0
Calculated life:
standard bearing L10= 92.2 hhigh-load capacity bearing L10= 120.4 h

Accumulated failure probability, %Fig. 3 Life test results under clean lubrication

Standard bearing
High-load capacity
bearing

95
90
80
70
60
50
40
30

slope e20

L (50)

1.971
3.275

22.5
49.0

10

L (10)

5

8.7
27.6
1
×100

2

3

5

7

1
×101

2

3

Life, hFig. 4 Life test results under contaminated lubrication

-32-

5

7

1
×102

2

High Capacity Tapered Roller Bearings

The results of the life test under the Λ=0.8 clean oil
lubrication condition are shown in Fig. 5.
Our high-load capacity tapered roller bearing wasless prone to develop surface-initiated flaking becauseof metal-to-metal contact due to poor oil film formation
conditions and boasts a bearing life approximately
twice as long as standard tapered roller bearings.
The results of the life test under the Λ=0.8
contaminated oil lubrication condition are shown in
Fig. 6.
We performed a market study of the contaminantsin lubricating oils and found the size of contaminants
was 50μm or smaller and that the amount ofcontaminants was less than 0.3 g/L even on MT cars.
Therefore, we employed severe contamination
conditions with contaminants sized 50μm or smallerin amounts of 0.3 g/L.
Our high-load capacity tapered roller bearing wasless prone to developing dent-initiated flaking during
use with contaminated lubricant and boasts a bearing
life approximately 3 times the length of the standard
tapered roller bearing.

5. Application to extremely low
torque/high rigidity design
Our high-load capacity tapered roller bearing can be
combined with a long-life bearing, such as the NTN FA
tapered roller bearing 2) that features a special heat
treatment process (FA), to obtain finer crystal grains
and optimized bearing interior design technologies.
The FA tapered roller bearing offers longer bearing
life, greater seizure resistance and improved dent
resistance. If it is designed for a bearing life equivalent
to standard tapered roller bearings, its size can be
more compact than that of standard bearings,
resulting in lower torque. However, loss in bearing
rigidity was unavoidable when we attempted toachieve a significantly lower torque. By combining the
FA tapered roller bearing with our newly developed
high-load capacity tapered roller bearing, it is possible
to avoid loss in bearing rigidity and achieve a lower
torque.

Accumulated failure probability, %Data for samples whose tests were terminated before they failed
Load conditions:radial load Fr= 19.1 kN
axial load: Fa= 0 kN
(axial clearance 0.20-0.25 mm)
Speed: 2500 min-1
Lubricant: Eneos Super Oil T10 (VG10)
Lubrication method: oil bath
Oil film parameter: Λ = 0.8
Calculated life:
standard bearing L10= 92.2 h
High-load capacity bearing L10= 120.4 hStandard bearing
High-load capacity
bearing

95
90
80
70
60
50
40
30

slope e20

L (50)

1.762
3.328

146.0
188.1

10
5

L (10)

7

1
×101

2

3

5

7

1
×102

2

3

5

50.1
106.8

7

Life, hLoad conditions:radial load: Fr= 19.1 kN
axial load: Fa= 0 kN
(axial clearance 0.20-0.25 mm)
Speed: 2500 min-1
Lubricant: Eneos Super Oil T10 (VG10)
Contaminants: steel beads + gas-atomization 0.3 g/LContaminant grain size
50μm or smaller, 90wt%
100-180μm, 10wt%
Lubrication method: oil bath
Oil film parameter: Λ= 0.8
Calculated life:
standard bearing L10= 92.2 hhigh-load capacity bearing L10= 120.4 h

Accumulated failure probability, %Fig. 5 Life test results under clean lubrication

Standard bearing
High-load capacity
bearing

95
90
80
70
60
50
40
30

slope e20

L (50)

1.685
1.158

49.8
237.2

10

L (10)

5

16.3
46.6
1
×100

5

1
×101

5

1
×102

Life, hFig. 6 Life test results under contaminated lubrication

-33-

5

1
×103

5

NTN TECHNICAL REVIEW No.73（2005）

Furthermore, it is possible to decrease the torque to
50% by greatly reducing the lubricant stirring drag inthe bearing through optimization of the cage shape.
Applications of 50% low torque design that also
ensure bearing life and rigidity are described below.

Based on this finding, we have summarized thetorque factors and contribution rates of a tapered roller
bearing used on a pinion shaft support in a near axle
differential in the pie chart below.
When the amount of lubricant on a bearing isrelatively large and its viscosity is relatively high, themajor contributing factors are "stirring torque caused
by lubricant viscosity," "rolling viscous torque on rolling
contact surface" and "shear torque on lubricant
between cage and rolling elements."

5.1 Tested bearing
Since the tapered roller bearings for pinion shaft
supports in near axle differentials are used under highloads, bearings capable of carrying large loads areused. Since the torque loss caused by the bearings isa relatively large part of the total loss, the need for
low-torque tapered roller bearings is growing.
To address this problem, we have attempted toachieve lower torque for a bearing used for pinion shaft
support in near axle differentials with model #30306D.

Rotational torque，N･cm

200

5.2 Torque factors on tapered roller bearings
The torque factors on tapered roller bearing areshown in Fig. 7.
To decrease the torque occurring on tapered roller
bearings, factors 1 through 5 need to be reduced.

1 Rolling viscous torque on
rolling contact surface

High oil levelLow oil level

160

30% torquereduction
120
80
40
0
0

1000

2000

3000

4000

5000

Bearing speed, r/min
Load conditions: axial load 6000 N
Lubricant: hypoid gear oil
Lubrication method: oil bath (natural circulation)

2 Sliding torque on rib surface

Fig. 8 Torque of tapered roller bearing used for pinion
shaft support in near axle differential

3 Torque resulting
from elastic deformation of rolling
elements
5 Stirring torque causedby viscosity of lubricant

High oil level

4 Shear torque on lubricant between
cage and rolling elements
Low oil level

Fig. 7 Torque factor of tapered roller bearing

5.3 Torque factors and their contribution totorque on tapered roller bearings for nearaxle differentials

Fig. 9 Torque measurement method

Among the torque factors working on a tapered
roller bearing, the magnitude of the stirring torquecaused by lubricant viscosity greatly varies dependingon the amount and temperature of the lubricant oil.
We executed tests with various oil levels, as shown in
Fig. 9, to determine the magnitude of stirring torqueon tapered roller bearings used for pinion shaft
support in near axle differentials. As a result, we havelearned that in a normal bearing running speed rangeof 2000-3000 r/min, the lubricant stirring torqueaccounts for approximately 30% of the whole torque.

Stirring torquecaused bylubricant viscosity

30%
57%

Shear torque on lubricant
between cage and rolling
elements

Rolling viscou
torque on rolling
contact surface

11%

Torque resulting from elastic 2%
deformation of rolling elements

Sliding torque on rib surface 0%

Fig. 10 Factors contributing to friction torque of tapered roller
bearing used for pinion shaft support in near axle differential-34-

High Capacity Tapered Roller Bearings

5.4 Torque reduction techniques30
Calculated values
Actual measurements

Reduction of the major contributing torque types isnecessary to decrease the torque on tapered roller
bearings used on pinion shaft support in near axle
differentials. These include "rolling viscous torque on
rolling contact surface", "shear torque on lubricant
between cage and rolling elements" and "stirring
torque caused by lubricant viscosity".

Torque reduction %

25
20
15
10
5

1 Decrease in rolling viscous torque on rolling
contact surface

0
0

1000

2000

3000

4000

5000

Bearing speed r/min

Optimization of the bearing interior design and thecrowning of rollers is effective for decreasing therolling viscous torque on a rolling contact surface. Our
torque reduction techniques that can ensure bearing
rigidity are described below.
1) Alteration of bearing internal design
To achieve a lower torque while maintaining
bearing rigidity, we have checked the bearing
internal design's contribution to torque and bearing
rigidity. The results are summarized in Fig. 11. To
achieve a high-rigidity low-torque design, smaller
roller pitches and greater contact angles areadvantageous.

Fig. 12 Torque reduce percentage by changing crowning

Torque ratio, rigidity ratio

1.3
1.2
1.1
0

0.5

1.0
1

1.5

2

Torque ratio
Rigidity ratio

0.9
0.8
0.7

Inner ring crowning radius ratio

Larger roller
diameter

Torque reduction effectHigh rigidity effectLonger roller
Larger rollerpitchLargercontact angle
Increased
number ofrollers-2

Crowning shape examples
-1.5

-1

-0.5

0

0.5

1

1.5

2

Fig. 13 Factors contributing to friction torque and rigidityof inner ring crowning

Degree of effectFig. 11 Effect of internal design factors in torque and rigidity

2) Alteration of crowning shape
Fig. 12 summarizes the torque reduction(calculated values and actual measurements) that
result from the 17% reduction in the effective
contact length between the rollers and the raceway
surface through alteration of the crowning shape.
The torque can be decreased by altering thecrowning shape to reduce the effective contact
length. The actual effect nearly matches our
calculated values.
However, as summarized in Fig. 13, a lower
torque attained by alteration of the crowning shape
can decrease the bearing rigidity. Therefore,
bearing rigidity must be considered to determine an

optimal crowning shape.

2 Reduction of shear torque on lubricant
between cage and rolling elements
As summarized in Table 2, compared to a standard
cage, the shape A cage with straight bars results in alower shear torque on the lubricant between the cageand rolling elements. The actual shear torquemeasurements match our calculations well. In thecase of shape B, the grooves on the ribs help promote
oil flow toward the outer ring. This arrangement canlower the lubricant stirring torque compared to cases
where lubricant oil tends to remain stagnant in the
bearing inner ring.
-35-

NTN TECHNICAL REVIEW No.73（2005）

Table. 2 Type of cage and torque reduction ratio
Standard cage

Shape A

Shape B

―

10%

12%

0%

10%

10%

Smallerdiameterside
Largerdiameterside
ActualTorquemeasurements
reduction
percentage Calculated
valuesOil flow

The torque reduction percentage resulting fromthe modified internal design is summarized in Fig.
16. Since the calculated values do not include thereduction in stirring torque caused by the low
viscosity of the lubricant, the reduction effect in thecalculated values is lower than the actual
measurements. However, when the reduction in the
lubricant stirring torque due to the roller rolling area
reduction is considered, as demonstrated byexpression (1), the calculated reduction in bearing
torque coincides well with the actual measurement.

3 Decrease in stirring torque caused by
viscosity of lubricant
1) Effect by modified bearing internal design
A low torque design with a more compact bearing
results in reduced roller rolling area volume asshown in Fig. 14, thereby reducing the lubricant
stirring torque inside the bearing. An example ofreduced roller rolling area volume is described
below.
Fig. 15 shows a comparison between thestandard bearing and a new bearing with a modified
internal design. As summarized in Table 3, therolling area volume in the bearing with a modified
internal design is 31% smaller than the standard
bearing.

Lubricant stirring
torque rate

Roller rolling area
volume reduction rate

Rolling viscosity
torque reduction rate

Bearing torque＝30%×（100%−31%）＋70%

×（100%−13%）＝82%（18% reduction）…Expression (1)Rolling viscosity
torque rate

Calculated value coincides well with actual measurement.

Table. 3 Comparison of bearing internal design
Roller length Avg. roller dia. Roller pitch dia.

Fig. 14 Roller rolling area volume
New bearing with modified
internal designStandard bearing
20.75

20.75

14

14

Roller rolling area
volume ratio

Standard bearing

13

9.22

51.54

1.00

Bearing withmodified internal
design

11

8.04

48.46

0.69

30
Calculated value
Actual measurement

Torque reduction %

φ72

φ30

19

φ72

φ30

25

19

20

18%

Approx. 5%

15

13%
10
5
0
0

1000

2000

3000

4000

5000

Bearing speed r/min

Fig. 16 Relationship between torque and torque reduce
percentage

Fig. 15 Comparison of bearing form

-36-

High Capacity Tapered Roller Bearings2) Alteration to cage inside diameter
To reduce the lubricant stirring drag by decreasing
the amount of lubricant flowing into the bearing, we
verified the torque reduction effect with a bearingthat has a small diameter cage. The resultant torquereduction effect is summarized in Fig. 17. At abearing speed of 2500 r/min, approximately 25%
reduction in torque is achieved. The lubricant
stirring torque accounts for approximately 30% of
the whole bearing torque. In other words, thelubricant stirring torque was reduced by 83%.

5.5 Ultra-low torque/high-rigidity tapered roller
bearing design example
By combining the FA bearing, the high-load capacity
design, and the torque reduction techniques, it ispossible to reduce bearing torque by 50% while
ensuring bearing life and rigidity (Fig. 18, Table 4).

Low-torque bearing

Standard bearing

16
20.75

10.4

19
Torque reduction %40

More compact
bearing design
is possible.

16

φ60

Small diameter cage

φ30

φ30

Standard cage

φ72

14

30

20

Fig. 18 Comparison of bearing size

10

Table 4 Bearing internal designStandard bearing

0
0

1000

2000

3000

4000

5000

Rated load

Bearing speed, r/min

15

17

Contact angle

28°48′39″

30°30′

Roller length (mm)

13

10.1

Mean roller dia. (mm)

φ9.22

φ7.25

Mass (kg)

0.393

0.223

1) Torque reduction effect by factor(calculated value)

Running torque reduction %550Fig. 19 summarizes the running torquereduction effect (expected value) for eachtorque reduction factor. In the practical
speed range of 2000 to 3000 r/min, it ispossible to reduce running torque by 50%.

4
3

40

2
30

1
20
1 Effect of modified internal design
2 1+crowning effect
3 2+effect of reduction in shear torque on lubricant between cage and rolling elements
4 3+effect of reduction in lubricant stirring toque by modified internal design
5 4+effect of reduction in lubricant stirring torque by limiting the amount of lubricant

0
0

1000

2000

3000

4000

5000

Bearing speed, r/min

Fig. 19 Relationship between speed and torque of each torque reduction factors-37-

φ44.44

Number of rollers

60

10

φ51.54

Roller PCD (mm)Fig. 17 Relationship between speed and torque reduce
percentage of bearing with small diameter cageLow-torque design

Dynamic rated load Cr =49.0 kN Dynamic rated load Cr =33.0 kN
Static rated load Cor=52.5 kN Static rated load Cor=35.5 kN

NTN TECHNICAL REVIEW No.73（2005）

2) Torque reduction effect (actual measurement)4) Axial rigidity measurement results

Fig. 20 summarizes the running torquemeasurement results for the standard and low-torquebearings.
In the practical speed range of 2000 to 3000 r/min,running torque was reduced by 50%.

Fig. 22 summarizes the axial rigidity measurement
results for the standard and low-torque bearings.
The rigidity of the low-torque bearing is equivalent
to that of the standard bearing.

60
50

Low-torque bearing
250

Axial elasticity μm

Running torque reduction %300

Standard bearing

200
150

50% reduction

40
30
20
Low-torque bearing
Standard bearing

10

100
0
0

5000

10000

0

15000

20000

25000

30000

Axial load N

50

Fig. 22 Deformation of axial direction
0

1000

2000

3000

4000

5000

Bearing speed r/min

Fig. 20 Relationship between speed and torque of current
bearing and low torque bearing5) Torque factor analysis for ultra-low
torque/high-rigidity tapered roller bearing
Fig. 23 summarizes the torque values obtained
from our verification efforts for the ultra-low torque,high-rigidity tapered roller bearing. The NTN ultra-low
torque, high-rigidity tapered roller bearing achieved
50% reduction in bearing torque while ensuring a levelof life and rigidity equivalent to that of the standard
bearing.

3) Reduction effect on running torque(calculated and actual values)

Running torque reduction %As shown in Fig. 21, the actual torque reductions
measured at various speeds match the corresponding
calculated values well.

60
Stirring torque from lubricant viscosity
50

Shear torque on lubricant between cage and rolling elements
Torque from elastic deformation of rolling elements

40

Sliding torque on rib surface (0%)
Rolling viscosity torque on rolling contact surface

30
20

30%
Expected value
Actual measurement

10
0

2%
0

1000

2000

3000

4000

11%

5000

Bearing speed r/min

5%
11%

0%

4%

Fig. 21 Relationship between speed and torque of low torquebearing

57%

0%
80%

Standard bearing

Low-torque design

Fig. 23 Analysis of torque reduction-38-

High Capacity Tapered Roller Bearings

6. Conclusion
As explained in this thesis, we believe that our highload capacity tapered roller bearing, which boasts
longer life and higher rigidity, offers a solution to
problems arising from challenges related to improving
automobile fuel efficiency. We will market long-life,
high-rigidity, low-torque tapered roller bearing
products that combine FA and high-load bearing
technologies to contribute to lower automobile fuelconsumption.

References
1) Tsutomu Ohki, Kikuo Maeda, Hirokazu Nakashima:
NTN Technical Review "Longer life with bearing steel
by finer crystal size," No. 71 (2003), p2.
2) NTN Catalog "FA Tapered Roller Bearings" (CAT. No.
3802/E）.

Photos of authors

Takashi TSUJIMOTO

Jiro MOCHIZUKI

Automotive Sales Headquaters
Automotive Engineering Dept.

Automotive Sales Headquaters
Automotive Engineering Dept.

-39-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Paper ]Study of Long-Life Thrust Needle Roller Bearings Used
in Low Viscosity Lubrication Conditions

Hiroki FUJIWARA**
Kenji TAMADA**This paper investigates the failure modes of thrust needle roller bearings lubricated with low
viscosity lubricants both experimentally and theoretically, and proposes a new design of a long-life
thrust needle roller bearing to meet these demands. Experimental observations reveal theoccurrence of surface originated flaking at the inner edge of the raceway, which suggests that heatgeneration due to sliding roller-race contact and the resulting material plastic flow were the likely
cause of the damage. Theoretical analysis shows that the dominant factor in bearing torque isslipping motion at the roller and raceway contact under insufficient lubrication film conditions.
Consequently, the use of crowned rollers and double rows of rollers are expected to yield longer life
as well as lower bearing rotational torque when compared with standard thrust needle roller bearings.

1. Introduction

Symbols

Thrust needle roller bearings, boasting various
advantages such as high-load carrying capacity, highrigidity and compactness, are used in a diversity of
applications. Their operating conditions andperformance requirements vary from application toapplication. Thrust needle roller bearings used for car
air-conditioner compressors use a fluid mixture
consisting of coolant and refrigerator oil as a lubricant.
Recently, lower viscosity fluids are being used to
improve the efficiency of compressors, and, as aresult, lubrication performance has been decreased.
To help solve this problem, long-life thrust needle
roller bearings capable of low viscosity lubrication are
increasingly needed.
The authors have investigated the factors that affectthe lifespans of thrust needle roller bearings under low
viscosity lubrication conditions and have attempted tofind measures to ensure longer bearing life.

b
dr
E'
EE
Fa
Fr
Ft
G
j
L
lrd
n
ME
Mr
Ms

**New Product Development R&D Center Mechatronics Research Dept.
**Intellectual Property Planning Dept.

-40-

Half width of contact
Roller diameter
Coefficient equivalent to Young's modulus
Elastic compression workload per time unit
Friction force occurring on protruded contact
area
Rolling viscous resistance
Traction force occurring on oil film
Material parameter
Roller row No.
Contact length
Length of rolling contact per time unit
Number of roller rows
Friction torque resulting from elastic
hysteresis loss
Frictional torque from rolling viscous
resistance
Frictional torque from spinning

Study of Long-Life Thrust Needle Roller Bearings Use in Low Viscosity Lubrication Conditions

P
R
U
v
W
w
w0
x
Z
α0
βE
δ
μ
ΦE
φ
φT
ω

Load
Equivalent radius
Velocity parameter
Absolute value of relative velocity between
race and rollers
Load parameter
Load per unit length
Load at minute width in contact center
Coordinates in roller revolution direction
relative to revolution center as origin
Number of rollers per row
Viscosity-pressure coefficient
Elastic hysteresis loss coefficient
Elasticity approach amount
Friction coefficient
Heat generation ratio resulting from elastic
hysteresis loss
Elastic compression workload per distanceunit
Compensation coefficient for rolling viscous
resistance resulting from shear heat generation
Angular velocity of turning ring

Bearings being tested

Hydraulicpressure

Driving motor

Fig. 1 NTN thrust needle roller bearing life test rig

Table 1 Life test condition(a) Life test operating conditions
Test rig

NTN thrust needle roller bearing life test rig

Load

9.8 kN

Bearing speed

5000 min-1

2.1 Life test method

LubricantSpindle Oil ISO VG2 (70˚C)

For this life test, the authors employed the NTN
thrust needle roller bearing life test rig. This rig isschematically illustrated in Fig. 1, and the test
conditions used are summarized in Table 1. The race
measures dia.φ60×dia.φ85×1 mm, and we usedvarious race materials and heat treatment furnaces.
The lubricant used was Spindle Oil ISO VG2. The oilfilm parameter (Λ) with this lubricant was 0.1 (surface
roughness 0.084 μmRa). Thus, the lubrication used
for the test was boundary lubrication.

Oil film parameter

0.1

Lubrication method

Circulating lubrication

2. Experimental investigation into
failure causes1)(b) Main characteristics of bearings used in life testRace

φ60×φ85×1 mm

Rollers

φ3×7.8 mm

20
18

SPCC

SCM415

Continuous carburizing
furnace

Batch furnace

Liquid carburizing
furnaceContinuous carburizing
furnace86420Batch furnace

10

Liquid carburizing
furnaceThree race materials, SPCC, SCM415 and SK5,
were used, and carburization/quenching was
performed with three different heat treatment furnaces.
The life test results with various races are shown in
Fig. 2. The end of bearing life for every bearing tested
was determined by flaking on the race. The life levelsof the races tested fell in a range of approximately 11
to 17 hours in terms of 10% life (L10), with no apparent
difference in the life resulting from race material orheat treatment furnace differences.

16
14
12

Continuous carburizing
furnaceL10 life, hours

2.2 Effects of race materials and heat treatment
furnaces

SK5

Fig. 2 Effect of race material and heat treatment furnace
on bearing life-41-

NTN TECHNICAL REVIEW No.73（2005）

whose grain boundary oxidizing zone was removed
with emery paper and buff-finished with diamondpaste was subjected to the life test, and the result was
as shown in Fig. 5.
The bearing life was not affected by the surface
roughness of the race, which was 0.062μmRa before
finishing and 0.027μmRa after buffing.
Furthermore, we investigated the possible
connection between lifespan and other factors,
including surface layer hardness, surface residual
stress and residual austenitic amount, and found no
significant relationships.

2.3 Effect of race accuracy
We investigated the effect of race surface
roughness on bearing life. We assumed that the race
surface roughness was in a range of approximately
0.03-0.11μmRa. The results are summarized in Fig.
3. No significant relation between surface roughness
and bearing life was found under these test
conditions.
Fig. 4 illustrates the results of investigation into theeffects of heat treatment on the deformation of theraces. The magnitude of heat treatment-induced
deformation was evaluated based on the flatness of
the races. Deformation had no apparent effect onbearing life.

50
40

L10 life in hours

L10 life in hours

50
40
30
20

30
20
10

10
0
0

0

0

0.02

0.04

0.06

0.08

0.10

0.12

7.3

Grain boundary zone thickness μm

Surface roughness (Ra) μm

Fig. 5 Effect of race grain boundary oxidizing zone on
bearing life

Fig. 3 Effect of race roughness on bearing life

2.5 Failure types
50

The appearance of the raceway that underwent thetest is shown in Photo 1. Most of the bearing failures
were shallow flakings that started at the inner bore
edges on the raceway rolling mark. Evidence of heatgeneration (discoloration) is apparent along the wholecircumference of the rolling mark edge. On thrustneedle roller bearings, rollers exhibited pure rolling on
the pitch circle with slip occurring towards both endsof each roller. The bearing life might have been
decreased by slip-induced heat generation.

L10 life in hours

40
30
20
10
0

31

51

102

103

Flatness μm

Fig. 4 Effect of race flatness on bearing life

2.4 Effects of race materials

Area A
Enlarged viewof area A

As shown in Fig. 2, variation of the race steel
material and heat treatment furnace type made nodifference on the life of the thrust needle roller
bearing. We believe the reason that no difference
occurred between the samples was that a grainboundary oxidizing zone that uniformly lowered thesurface strength was present on the surface of eachrace. The depth of the grain boundary oxidizing zone
was in a range of 1.5 to 9.4μm. To determine theeffect of the grain boundary oxidizing zone, a race

Flaking

Rolling markEnlarged viewof area B
Area B

Photo 1 Raceway after life test

-42-

Study of Long-Life Thrust Needle Roller Bearings Use in Low Viscosity Lubrication Conditions

Furthermore, the surface characteristics of the raceon the crowned roller bearing were compared to thoseof the straight roller bearing. Fig. 8 summarizes theresults of X-ray half-value width measurements on theinside and outside edges and the middles of the rollingmarks on the races of both bearing types that we
tested. Compared to the crowned rollers, the decreasein X-ray half-value widths at the edges of the straightrollers was greater. We found that there was greater
heat generation at the edges of the rolling mark on therace of the straight roller bearing compared to thecrowned roller bearing.
Fig. 9 illustrates the results of X-ray anisotropy
measurements on the inside and outside edges andthe middles of the rolling marks on the races of bothtested bearing types. "X-ray anisotropy" can be
defined as a phenomenon where the crystals areoriented in a particular direction due to the plastic
fluidization of the surface due to tangential force2).
This can be qualitatively described by using theabsolute value of the difference between the residual
stress measured on the rolling mark surface in theroller rolling direction and the residual stress

These tested rollers and races did not use crowning,and edge loading was considered one factor for
failure. If the edge loads were a major failure factor,
the failure should be internally started flaking.
However, the failure that occurred in our test was
flaking that started from the surface. In addition, assummarized in Fig. 6, the inner edge of the rollingmark on the raceway that underwent the testdeveloped virtually no compressive residual stress,
while the compressive residual stress on the outeredge reached a depth of 0.03 mm. The estimatededge stress was 2 GPa maximum, so we do not thinkthat the edge load is a major contributing factor.

Distance from surface mm
Residual stress, GPa

0

0

0.01

0.02

0.03

0.04

0.05

0.06

-100
-200
-300
-400

Inner edge

-500

Middle-600

Outer edge

-700

Straight roller

Fig. 6 Compressive residual stress in a race after test

Crowned roller

2.6 Experimental verification of assumed failurecauses
To verify our assumptions described above, we
introduced crowned rollers into the bearing being
tested so that the actual roller contact length was
shorter than with straight rollers and subjected thebearing samples to the life test. As a result, thecrowned roller bearing life was 4.5 times longer thanthat of the straight roller bearing, as shown in Fig. 7,
though its surface pressure was higher. From thisfinding, the relative slip between the rollers and therace seems to affect the bearing life.

MiddleOutside edge
New roller
Inside edge
MiddleOutside edge
New roller
5

6

6.5

7

Straight roller

Fig. 8 X-ray half-value width of rolling marks on races

80

Crowned roller

60
40
20
0

Inside edge
MiddleOutside edge
New roller
Inside edge
MiddleOutside edge
New roller0Straight roller

5.5

X-ray half-value width deg

100

L10 life in hours

Inside edge

Crowned roller

50

100

150

200

250

X-ray anisotropy MPa

Fig. 7 Comparing bearing life with straight rollers and one
with crowned rollers

Fig. 9 X-ray anisotropy of rolling marks on races

-43-

300

NTN TECHNICAL REVIEW No.73（2005）

measured in the opposite direction (for convenience,the estimated stress was determined with a simplified
sin2μmethod). The greater this difference, the greater
the plastic flow occurring on the surface.
The residual stress on the inner edge of the rollingmark on the straight roller bearing was approximately
160 MPa greater than that of the crowned roller
bearing.
From these facts, we believe that the relative slip
between the rollers and the race led to heatgeneration and plastic fluidity at the inner edge of therolling mark on the race, resulting in the surfaceinitiated flaking.

The friction coefficientμ(x) can be given by a mixed
lubrication theory. The frictional force on a solid-tosolid contact area was calculated using theGreenwood-Tripp theory 4), and the traction force on
the fluid lubrication area was calculated using MurakiKimura's analysis method. 5)
Incidentally, it is usually difficult to correctly provide
various parameters that represent the surface profileand lubrication state of a solid-to-solid contact model.
Therefore, with a thinner oil film and a greater
proportion of solid-to-solid contact, friction coefficientsand frictional torques do not necessarily coincidequantitatively with experiment results.

3.1.2 Rolling viscous resistance3. Analysis of frictional torque3)

According to Zhou-Hoeprich 6), shear stress on
lubricant at the EHL inlet opposite the rolling direction
occurs on the surface of the contact object. If
converted into a surface integral, this shear stress canbe evaluated as a force that causes a moment
opposing the revolution to occur. This force is known
as "viscous resistance." As a result of analysis of the
EHL theory, Zhou-Hoeprich developed a recurrence
formula for rolling viscous resistance, as shown in
expression (2), which we adopted.

The amount of heat generation can be given as the
product of frictional torque and bearing speed. Given
this, we will hereafter investigate the frictional torquefactors and their impacts theoretically.

3.1 Frictional torque factors and methods fortheir calculation
3.1.1 SpinAn example of the distribution of the peripheral
speed and the slip between the rollers and the race ona thrust needle roller bearing is illustrated in Fig. 10. Ifcrowning is ignored, the peripheral speeds within thecontact surfaces are uniformly distributed in therevolution axis. At the same time, the race velocity
distribution is proportional to its radius. Accordingly,
within the contact surfaces, slipping occurs in thenormal direction on the contact center, and thisslipping motion is known as "spin." When the friction
coefficient is μ(x), the bearing torque Ms caused byspin can be described in the form of the expression(1).

R
Fr＝φT 29.2 ――L (GU ) 0.648 W 0.246⋯⋯⋯⋯(2)α0

For thrust needle roller bearings, the frictional
torque Mr resulting from rolling viscous resistance canbe expressed as:nΣ∫x Fr (x) dx ⋯⋯⋯⋯⋯⋯⋯⋯(3)

Mr＝2Z

j=1

3.1.3 Elastic hysteresis loss
When elastic bodies repeatedly come into mutualcontact, it is known that the relation between the
resultant stress and strain can be plotted as ahysteresis loop. We determined the elastic hysteresis
loss for a line-contact case.
In the contact area in Fig. 11, the load on theminute width dy at the center of the contact area is:

2Z n
Ms＝―
―
ω

Σ∫μ(x) w(x) v(x) dx ⋯⋯⋯⋯(1)
j=1

Roller peripheral
speed

Race peripheral
speed

2P
1 E'LP 1/2
w0＝―― dy＝ ―― ――――
dy……………(4)
πb
2π R

（

RollerRace）

Therefore, the elastic compression workload dφconsumed when a roller rolls by dy is:

Slip on
race

1.4

dδ
P
― dP＝0.65 ――――――― dy ⋯(5)
∫w ―dPR E' L
P

dφ＝

0

0

Fig. 10 Peripheral speeds and slip speed distribution
in thrust needle roller bearing

-44-

0.5

0.4

0.3

Study of Long-Life Thrust Needle Roller Bearings Use in Low Viscosity Lubrication Conditions

Table 2 Frictional torque test condition(a) Operating conditions for friction torque measurement

L

Test rig

NTN vertical torque test rig

Load

3.0 kN

Bearing speed

1000〜3000 min-1
Spindle oil ISO VG2 (31˚C)

Lubricantdy

Turbine oil ISO VG32 (37˚C)
2bLubrication method

Contact area

1mR coating(b) Main data for frictional torque measurement test bearing

Fig. 11 Roller/race contact part

φ60×φ 85×1 mm

Race

φ3×7.8 mm

The rolling contact distance lrd per time unit wherea roller remains in contact with a race is:Single row
rollersω dr
lrd＝―― ―― …………………………………(6)2 2Crowned roller（φ3×7.8）

φ3×3.8 mm
Crowned roller（φ3×3.8）

Double row
rollers

Therefore, the elastic compression workload EE is:φωdr
EE＝φlrd ＝――― ……………………………(7)4Then, if the elastic hysteresis loss coefficient istaken as βE, the relation between the heat generation
and bearing frictional torque resulting from the elastic
hysteresis loss is:

Driving motornΦE＝2ZβE

ΣEE＝ωME ⋯⋯⋯⋯⋯⋯⋯⋯(8)

Bearing tested
Load cell(frictional force
detection)

j=1

Air staticpressure table

Therefore, the frictional torque ME resulting from theelastic hysteresis loss can be determined as:

Σ

2ZβE
EE
j=1
ME＝――――――⋯⋯⋯⋯⋯⋯⋯⋯⋯⋯⋯(9)ωLoad

Air cylinder

3.2 Comparison between measured frictional
torques and calculated results

Fig. 12 NTN vertical type bearing frictional torque test rig

Table 2 summarizes the test conditions for thefrictional torque measurement test as well as the
major technical data for the bearings tested withcrowned rollers. The temperature was measured at
the rear face of the race. The test rig used isschematically illustrated in Fig. 12.
In both experimental and calculated results with a
VG2 lubricant, plotted in Fig. 13, the frictional torquedecreases as the bearing speed increases and thefrictional torque attained was lower with the single row
bearing than with the double row bearing. The
resultant oil film parameter fell in a range of Λ=0.7 to
1.6.

Frictional torque Nm

0.5

Experiment (single row)

Experiment (double row)

Calculation (single row)

Calculation (double row)

0.4
0.3
0.2
0.1
0
0

500

1000

1500

2000

2500

3000

3500

Bearing speed min-1

Fig. 13 Test and calculation results of bearing frictional
torque with ISO VG 2 oil-45-

NTN TECHNICAL REVIEW No.73（2005）

The results obtained with VG32 are summarized inFig. 14. The frictional torque remained unchanged orincreased in the experiment, while the frictional
torque increased in the calculation. There was noapparent difference between the single row bearing
and the double row bearing in either the experiment or
the calculation. The Λ fell in a range of 4.4 to 9.6.
With either condition, the value obtained from theexperiment was greater than that obtained from thecalculation. One possible reason for this is that thefrictional torque resulting from cage/roller contact was
not considered in our calculations. In addition, theresults from VG2 shown in Fig. 13 could not correctly
reproduce the parameter of solid-to-solid contact.

Frictional torque Nm

0.5

Experiment (single row)

Experiment (double row)

Calculation (single row)

Calculation (double row)

When Λ=9, the major frictional torque factor isrolling viscous resistance, which accounts for 80% ormore of the total frictional torque. Spin does not
greatly affect rolling viscous resistance. As a result,
there is virtually no difference between a single row
bearing and a double row bearing, both having thesame effective contact length within a bearing.
When Λ=1, the lubrication condition is mixed
lubrication and the solid-to-solid contact area
increases. As a result, spin becomes the major
frictional torque factor. Consequently, compared witha single row bearing, the torque is lower on a double
row bearing, which has low spin.
The estimated temperatures on raceway surfaces
obtained through FEM are summarized in Fig. 16. The
temperature on the inner edge is higher than on theouter edge and the possibility of failure is higher with
the inner edge. In addition, it is apparent that the useof crowned rollers makes the temperature lower.
In summary, we have verified theoretically that
torque on a thrust needle roller bearing in mixed
lubrication conditions can be decreased by introducing
double row crowned rollers. A lower torque in turnhelps lower the temperature on the contact surface
and prevents bearing material failure.

0.4
0.3
0.2
0.1
0
0

500

1000

1500

2000

2500

3000

3500

Bearing speed min-1

Fig. 14 Test and calculation results of bearing frictional
torque with ISO VG 32 oil
Temperature ˚C

3.3 Analysis and discussion of torque factors

0.5

0.2

0.1

80

Effective roller
contact area

60
2

4

6

8

Elastic hysteresis loss
Spin

Frictional torque Nm

Frictional torque Nm

0.3

100

10

12

14

Fig. 16 Estimated temperature on raceway by FEM

Spin
0.4

120

Distance from bore or race mm

Elastic hysteresis loss

Rolling viscous resistance

140

40 0

The breakdowns of frictional torque, with a larger oilfilm parameter Λ and a smaller Λ are summarized inFig. 15.

0.5

Straight roller
Crowned roller

160

0.4

Rolling viscous resistance

0.3

0.2

0.1

0

0

Single row
Double row(a)Λ = 9

Single row
Double row(b)Λ = 1

Fig. 15 Oil film parameter and factors of bearing frictional torque-46-

Study of Long-Life Thrust Needle Roller Bearings Use in Low Viscosity Lubrication Conditions

References

4. Conclusion

1) Kenji Tamada, Factors affecting the life of thrustneedle bearings, Tribology Conference Preprints(Tokyo 2002-5), 131.
2) V.M.Hauk, Evaluation of macro- and micro-residual
stresses on textured materials by X-ray, neutron
diffraction and deflection. measurements, Adv. X-ray
Anal., 29(1986), 1.
3) Hiroki Fujiwara, Kenji Fujii, Torque analysis for double
row thrust needle roller bearings, The Japan Society
for Precision Engineering 2003 Spring Meeting
Scientific Lectures Proceedings, (2003), 585.
4) J.A.Greenwood and J.H.Tripp, The Contact of Two
Nominally Flat Rough Surfaces, Proc. Inst. Mech.
Eng., Tribol., 185, 48/71(1970-71), 625.
5) Masayoshi Muraki, Yoshitugu Kimura, Research into
traction characteristics of lubricants (2nd report)"Thermal analysis of traction with nonlinear
viscoelasticity", Junkatsu, 28 (1983), 753.
6) R.S.Zhou and M.R.Hoeprich, Torque of tapered roller
bearings, Trans. ASME, J. Tribol., 113(1991), 590.

The authors experimentally and theoreticallyinvestigated the failure mechanism of thrust needle
roller bearings in low viscosity lubrication conditions.
From the results of various tests, we have verified thatheat generation on the inner edge of the raceway
rolling mark and the resultant plastic fluidity of the
raceway surface cause surface-initiated flaking onthese bearings. We have learned that under a low
viscosity lubrication condition, spin is largely
responsible for the frictional torque on a thrust needle
roller bearing and that spin-induced slip can be
decreased and excessive heat generation inhibited byintroducing crowned rollers and a double row
arrangement. Low heat generation helps suppress
plastic fluidity of the raceway, leading to a longer
bearing life.
The double row roller arrangement is described indetail in the article "Double Row Thrust Needle
Bearings" in this edition of the NTN Technical Review.

Photos of authors

Hiroki FUJIWARA

Kenji TAMADATechnical Research Dept.
Research & Development Center

Intellectual Property
Planning Dept.

-47-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Double-Row Thrust Needle Roller Bearings
for Automotive Air Conditioners and Automatic Transmissions

Kosuke OBAYASHI*

In order to meet the customer's need for longer service life, lower torque and lower noise bearings,
NTN has developed double-row thrust needle roller bearings.
By using two rows of needle rollers in each retainer pocket and by optimizing the roller shape,
NTN has successfully minimized roller slippage, achieved extended service lives and reduced torqueand noise in thrust bearings.

1 Introduction

2. Features of double-row thrust needle
roller bearings (for automotive airconditioner compressors)

In recent automotive bearing applications, smaller
amounts of lower viscosity lubricants are being used
to achieve better fuel consumption. As a result, theoperating conditions for bearings have become
increasingly severe.
In this context, a unique problem for thrust needle
roller bearings have decreased bearing life because ofroller slipping. Bearings with longer lifespans areneeded.
We recently developed "double-row thrust needle
roller bearings" to cope with this market need. These
bearings have a double-row roller arrangement toinhibit the slipping of needle rollers and to achieve
longer life through prevention of surface-start failure.
They also have lower torque and, through roller shape
optimization, lower noise level.
For automotive air-conditioner compressors, wehave developed a resin cage to achieve lighter weight
and lower noise. For automatic transmissions, we
designed a new sheet steel cage with inner and outercircumferences that are secured by a full enclosuretechnique to prevent cage loosening and to enhance
cage mechanical strength. These cage types aredescribed below.

The thrust bearings for automotive air-conditioner
compressors are run with limited amounts of lowviscosity lubricant in order to improve the capability of
the compressors. Lower viscosity lubrication will befurther needed to help realize better automobile fuelefficiency. Reduced bearing life poses a problem to besolved.
In addition, as CFC refrigerant is replaced with CO2
refrigerant to help preserve the global environment, ahigher refrigerant compression ratio is increasingly
needed, which in turn leads to decreased bearing life.
Our newly developed double-row thrust needle
roller bearing (Photo 1) maintains load carrying
capacity and suppresses roller slipping. The doublerow roller arrangement prevents surface-start failuresthat result from disrupted oil film on bearing rollingsurfaces. Our new bearing adopts optimal roller
geometry (optimal end face shape, roller circularity
and crowning) and a lightweight resin cage (of highwear resistant material) to achieve longer bearing life,
lower torque and lower noise.

*Automotive Sales Headquarters Needle Roller Bearings Engineering Dept.

-48-

Double-Row Thrust Needle Roller Bearings

Photo 2 Current bearings

Photo 1 Double-row thrust needle roller bearingsfor car air conditioners[Comparison with single-row bearing in termsof performance]

2.1.1 Bearing specification
The bearings used for the performance evaluation
have 60 mm inside diameters, 85 mm outside
diameters and are 3 mm wide (roller diameter).
Table 1 lists the data for both our newly developed
bearing and the standard bearing.

(NTN's test results for automotive air-conditioner
compressors)
1 Life (surface-starting failure) : 12 times as long
2 Rotation torque: 50% reduction
3 Noise: 8 dBA lower

Table 1 Comparison of bearing specifications
Newly developed bearing(double-row type)

2.1 Performance evaluation
To evaluate performance of thrust needle roller
bearings for automotive air-conditioner compressors,
we compared our newly developed bearing with astandard bearing that has a W-type sheet steel cage
and a single-row roller arrangement (Photo 2). We
subjected both bearing types to rotation torque andrunning noise measurements after a bearing life test.
This section summarizes the performance evaluation
tests of both bearing types.

Bearing size
Cage type

Standard bearing(single-row type)

φ60ID×φ85OD ×3W (roller diameter)
Resin cage

W-type sheet steel cage

Number of rollers

96 (48 ×2 rows)

48

Roller length (mm)

3.8

7.8

2.1.2 Bearing life test
For the bearing life test, we used the NTN verticalthrust bearing test rig illustrated in Fig. 1 and the NTN
horizontal thrust bearing test rig shown in Fig. 2.

Load

Bearing being tested

Bearing being tested

Load

Fig 1 NTN vertical type thrust bearing test rig

Fig 2 NTN horizontal type thrust bearing test rig-49-

NTN TECHNICAL REVIEW No.73（2005）

Accumulated failure probability, %(1) Test conditions
Two sets of bearing life test conditions are given in
Tables 2 and 3.
Table 2 Life test condition 1
Test rig

NTN vertical thrust load test rig

Axial load

P/C = 0.4 with standard bearing＊

Bearing speed

5000 min-1

Misalignment on race

2/1000

Lubrication

Turbine #32, circulating lubrication

Number of samples

n=8, each bearing type

95
90
80
70
60
50
40
30
20

Standard bearing
Newly developed
bearing

10
5

5 7

1
101

2

3

5 7 1
102

2

3

5 7 1
103

2

3

5

7

Life, h＊P/C : load/basic dynamic rated load

Fig 3 Result of life test condition 1 with misalignment
for car air conditioners

Test rig

NTN horizontal type thrust load test rig

Axial load

P/C = 0.4 with standard bearing

Bearing speed

5000 min-1

Lubrication

PAG (3%) + clean kerosene (97%),
circulating lubrication

Number of samples

n=8, each bearing type

Accumulated failure probability, %Table 3 Life test condition 2(2) Test results
The life test result of test condition set 1 issummarized in Fig. 3, and that from test condition set
2 is given in Fig. 4.
Compared with the standard bearing, our newly
developed bearing exhibited a life more than 12 times
as long in terms of L10 under both test condition sets 1
and 2.

95
90
80
70
60
50
40
30
20

Standard bearing
Newly developed
bearing

10
5

1
101

2

3

5 7

1
102

2

3

5

7

1
103

2

3

5

Life, hFig 4 Result of life test condition 2 with lean lubrication
for car air conditioners

Rotation

2.1.3 Rotation torque measurement
For measuring the rotation torque of the bearings
being tested, the NTN vertical thrust bearing rotation
torque test rig was used.
The test rig is illustrated in Fig. 5

Bearing being
tested(1) Measuring conditions
Rotation torquedetection load cell

The rotation torque measuring conditions are listed
in Table 4.

Table 4 Measurement condition of rotation torque
Test rig

NTN vertical-type thrust torque test rig

Axial load

P/C = 0.1, 0.15, 0.2 with standard bearing

Bearing speed

1000 min-1

Lubrication

Category 1 spindle oil, overall coating

Load

Fig 5 NTN vertical type thrust bearing rotation torquetest rig

Measuring duration 60 sec.
Number of samples n=10, each bearing type-50-

Double-Row Thrust Needle Roller Bearings(2) Measurement results

2.1.4 Sound measurement

The measurement results of rotation torque areplotted in Fig. 6.
These results verify that the newly developed
bearing has achieved 47 to 51% torque reduction
compared with the standard bearing.
This torque reduction was achieved by both the
double-row roller arrangement, which reduced spin
slippage, and the cage resin material.
The measured rotation torques were dependent on
loading, but rotation speed caused virtually novariation.

For sound measurement, the NTN sound levelmeasuring system was used.

(1) Measuring conditions
The sound level measurement conditions are
summarized in Table 5.

(2) Measurement results
The sound level measurement results are given in
Fig. 7.
Compared with the standard bearing, the newly
developed bearing achieved a sound level reductionof approximately 8 dBA at 1800 min-1, andapproximately 11 dBA at 3600 min-1.
The contributing factors for these reductions are theresin-made cage and the optimized roller shape.

Standard bearing: W-type cage + single-row rollers
Newly developed bearing: resin cage + double-row rollers

60

Table 5 Measurement condition of sound level
40

20

0
0.05

0.10

0.15

0.20

0.25

P/C

Fig 6 Measurement result of rotation torque

Test rig

NTN sound level measuring system

Axial load

100N

Measurement distance45˚ angle, 100ｍｍ

Bearing speed

1800 min-1，3600 min-1

Lubrication

Compressor oil,droplet lubrication before running

Measuring duration

50 sec.

Number of samples

n=10, each bearing type

Standard bearing: W-type cage + single-row rollers
Newly developed bearing: resin cage + double-row rollers
90Sound level dBA

Rotation torque N・cm

80

85

80

75

70

65

1800 min-1

3600 min-1

Fig 7 Measurement result of sound level-51-

NTN TECHNICAL REVIEW No.73（2005）[Comparison with single-row bearing]

3. Feature of double-row thrust needle
roller bearings(for automatic transmissions)(NTN's test results assuming application in
automatic transmissions)
1 Life (surface-starting failure): Three times as long
2 Rotation torque: 40% decrease
3 Cage strength: 1.6 times as strong
4 Sound level: Decrease of 3 dB

Featuring a lower profile in assembled units and agreater load carrying capacity, thrust needle roller
bearings are often used for automatic transmissions in
automobiles. However, the use of lower viscosity
automatic transmission oils is increasing to alleviate
shift shock and improve fuel efficiently. For thisreason, the life of bearings used in this application
may be shortened.
Our newly developed double-row thrust needle
roller bearing (Photo 3) for automatic transmissions
has a double-row roller arrangement to inhibit rollerslipping while maintaining the necessary load carrying
capacity. The newly developed bearing features
optimized roller geometry (with optimized roller end
face shape, roller roundness and crowning shape) toachieve a longer bearing life, lower rotation torque andquieter operation.
In addition, because thrust bearings are used in
eccentric rotation conditions in automatictransmissions, cages can be pinched and damaged bybearing rings and the cage halves themselves cansplit apart. Therefore, higher cage strength is needed.
To address this problem, we have invented anddeveloped a new cage uniquely shaped with full
enclosure of inner and outer ring circumferences,improving cage mechanical strength. This
arrangement prevents separation of cage halves ordisplacement between the halves and allows
integration of the cage with the bearing ring to form aunified structure.

3.1 Performance evaluation
Under conditions that simulated automatictransmissions, the newly developed double-row thrustneedle roller bearing was compared with a standard
bearing that has a W-type sheet steel cage andsingle-row rollers. Performance evaluations wereconducted through FEM analysis, bearing life tests,
eccentric rotation tests, rotation torque measurements
and sound level measurements. The results are
summarized below.

3.1.1 Bearing specifications
The bearings used for performance specification
were 60 mm inside diameter, 85 mm outside diameterand are 3 mm wide (roller diameter).
The bearing specifications for the newly developed
bearing and the standard bearing are summarized inTable 6.

Table 6 Comparison of bearing specifications
Newly developed bearing(double-row type)
Bearing size

φ60ID×φ85OD ×3W (roller diameter)

Cage type

Sheet steel cage of new shape W-type sheet steel cage

Number of rollers

96 (48 ×2 rows)

48

Roller length (mm)

3.8

7.8

Double-row rollers

Full circumference enclosure

Photo 3 Double-row thrust needle roller bearings for automatic transmissions-52-

Standard bearing(single-row type)Double-Row Thrust Needle Roller Bearings

The maximum principle stress on the standard
bearing was compared with that on the newly
developed bearing. As a result, the stress on the
newly developed bearing was 0.62 relative to that of
the standard bearing rating of 1. The cage of the
newly developed bearing has increased strength.
Note: The weight of the newly developed bearing is
5% lighter than the standard bearing.

3.1.2 FEM analysis
The results of FEM analysis for the newly
developed bearing and the standard bearing aresummarized below.

(1) Analysis conditions
The analysis conditions are shown in Fig. 8.

(2) Analysis results
The analysis results are given in Fig. 9.

Fully constrained(inner circumference surface)

Load (50% load on the outercircumference surface and
50% load on the bars andpockets)

Fig 8 Analytical condition
Standard bearing

Newly developed bearing

Location where maximum principle load occurs

Fig 9 Analytical result

Accumulated failure probability, %3.1.3 Bearing life test
For the bearing life test, the NTN vertical thrust test
rig illustrated in Fig. 1 was used.

(1) Test conditions
The test conditions are summarized in Table 7.

(2) Test results
The results of bearing life tests are summarized inFig. 10.
In L10 terms, the life of the newly developed bearing
was three times as long as that of the standard
bearing.

NTN vertical type thrust test rig

Axial load

P／C=0.4 on standard bearing

Bearing speed

8000 min-1

Lubrication

ATF, circulating lubrication
Oil temperature, 100˚C

Number of samples tested

n=8 eachStandard bearing
Newly developed
bearing

10
5

5

7

1
101

2

3

5

7

1
102

2

Life, hTable 7 Life test condition
Test rig

95
90
80
70
60
50
40
30
20

Fig 10 Result of life test for automatic transmissions-53-

NTN TECHNICAL REVIEW No.73（2005）

For the eccentric rotation test, the NTN eccentricrotation vertical thrust test rig was used.
The test rig used is illustrated in Fig. 11.

3.1.4 Eccentric rotation test
If the inner ring and outer ring rotate on an offset
rotation center and the eccentricity is greater than theclearance of the bore, then the cage will be pinchedbetween the races (inner ring and outer ring).
Consequently, the cage is subjected to a compression
force in the radial direction and a shear force in therotation direction resulting from the difference between
the rotation speed of the inner race and that of theouter race, leading to eventual fracture.
The situation above exists particularly in the torqueconverters of automatic transmissions. To address
this problem, we checked the mechanical strength of
the new cage shape.

(1) Test conditions
The test conditions for the eccentric rotation test aresummarized in Table 8.

(2) Test results
The results of the eccentric rotation test are given in
Table 9.
A ring-shaped fracture occurred on the bore side of
the cage bar of the conventional bearing. In contrast,
no fracture was found in the newly developed cage.
Compared with the conventional cage, the newcage, with a unique form created by full enclosure of
inner and outer circumferences, exhibited improved
strength that matches the results of FEM analysis.

Load
Bearing
tested

3.1.5 Rotation torque measurement
The rotation torque of the bearings was measured
with the NTN vertical thrust torque test rig illustrated inFig. 5.

(1) Measuring conditions
The rotation torque measuring conditions are
summarized in Table 10.

(2) Measurement results
The rotation torque measurement results areillustrated in Fig. 12.

Table 10 Measurement condition of rotation torqueFig 11 NTN vertical type thrust bearing test rig
for eccentricity rotation

Table 8 Eccentricity rotation test condition
NTN eccentric rotation vertical

Test rig

thrust test rig

Axial load

P／C=0.3 on standard bearing

Bearing speed

3000 min-1

Eccentricity (rotation axis)

0.5mm

Lubrication

ATF, circulating lubrication
Oil temperature, 100˚C

Number of samples tested

n=3 each

Test rig

NTN vertical thrust torque test rig

Axial load

P/C = 0.1, 0.15, 0.2 with standard bearing

Bearing speed

1000 min-1

Lubrication

Category 1 spindle oil, overall coating

Measuring duration 60 sec.
Number of samples n=10, each bearing type

Standard bearing: W-type cage + single-row rollers
Newly developed bearing: resin cage + double-row rollers

Rotation torque N・cm

80

Bearing
No.

Running time, hRemarks

Standard
bearing

1

53.4

Ring-shaped fractureon the cage bar bore side

2

51.5

↑

3

60.5

↑

Newly developed
bearing

Table 9 Result of eccentricity rotation test

1

Suspended at 265.0 hrs.

No failure

2

Suspended at 265.0 hrs.

↑

3

Suspended at 265.0 hrs.

↑

60

40

20

0
0.05

0.10

0.15

0.20

P/C

Fig 12 Measurement result of rotation torque

-54-

0.25

Double-Row Thrust Needle Roller Bearings

As a result of the measurement, we found thatcompared with the standard bearing, the rotational
torque of the newly developed bearing is 38-43%
lower.
The primary factor contributing to the torquereduction is the reduction of spin slipping due to the
double-row roller design.
When comparing the newly developed cage with theconventional cage, other possible contributing factors
are a smaller contact area between the cage and shaft
and a smaller contact width between the cage androllers.
We also learned that the rotation torque isdependent on the load applied to the bearing tested
and is least affected by the bearing running speed.

Standard bearing: W-type cage + single-row rollers
Newly developed bearing: resin cage + double-row rollers

Sound level dBA

90

85

80

75

70

65

1800 min-1

3600 min-1

Fig 13 Measurement result of sound level

3.1.6 Sound level measurement
For sound level measurement, the NTN sound level
test system was used.

4. Conclusion(1) Measuring conditions

This report described our newly developed doublerow thrust needle roller bearings.
In this report, the scope of application for our newly
developed bearings were limited to automobile airconditioner compressors and automatic transmissions
and we evaluated them only for these applications.
However, these new bearings are also capable ofapplications in other types of machinery.
The operating conditions for bearings will become
increasingly severe. Therefore, we will remaincommitted to the development and marketing of novel
bearing products that only NTN can offer, always
aiming to increase customer satisfaction and satisfy
market needs.

The sound level measuring conditions are
summarized in Table 11.

(2) Measurement results
The sound level measurement results obtained areillustrated in Fig. 13.
At 1800 min-1, the sound level of the newly
developed bearing was approximately 3 dBA lower,
and at 3600 min-1, approximately 2 dBA lower.
The factor contributing to this sound suppression is
optimization of the roller geometry.
Table 11 Measurement condition of sound level
Test rig

NTN sound level measuring system

Axial load

100N

Measurement distance45˚ angle, 100ｍｍ

Bearing speed

1800 min-1，3600 min-1

Lubrication

ATF,
droplet lubrication before operation

Measuring duration

50 sec.

Number of samples

n=10, each bearing type

Photos of author

Kosuke OBAYASHI
Automotive Sales Headquraters
Needle Roller Bearings
Engineering Dept.

-55-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Article ]Development of In-Wheel Type Integrated Motor Axle Unit

Minoru SUZUKI*
Dawei WANG*

NTN has been developing the in-wheel type integrated motor axle unit for electric vehicles. The unit
accomplishes compact size and light weight by integrating a hub bearing, a cycloid reducer and ahigh-speed motor, realizing least increase in the unsprung mass of in-wheel motor vehicles. This
paper introduces the design concept and basic performance of the unit.

concept and initial characteristics of this novel axle
unit.

1. Introduction
Recently, fuel-cell electric vehicles (FCEV) andelectric vehicles (EV) are being regarded as promisingnext generation automobiles because of their reduced
impacts on the environment. The drive systems for
EVs, including FCEVs, can be categorized into two
types. One type consists of a motor situated on the
chassis that transmits driving power to the tires
through a drive train, like a conventional internal
combustion engine. The other type is an in-wheel
motor system in which a motor is incorporated into
each wheel. On the positive side, an in-wheel motor
system allows the provision of room space equivalent
to or better than that possible in engine vehicles andhelps improve vehicle stability because each wheel isdriven independently. However, on the negative side,in-wheel motor system lead to greater unsprung massthat in turn results in decreased drivability and riding
comfort. Therefore, this issue needs to be resolved sothat EVs with in-wheel motor axle units can be morecommonly used. NTN has been developing an inwheel type integrated motor axle unit, aiming to
reduce unsprung mass by creating a compacter,lighter axle unit. This report describes the design

2. Newly developed axle unit specifications
We are currently developing an in-wheel type
integrated motor axle unit that, when built into four
wheels, has power equivalent to the 1500 cc gasoline
engine of a Japanese compact car. The target
specifications of our axle unit are summarized inTable 1. The structure of the axle unit is schematicallyillustrated in Fig. 1, and an image of the axle unit built
into a tire is shown in Fig. 2.
Table 1 Target specification of axle unit
Max. output

20kW

Max. torque490Nm

Max. velocity

150km/h (when built into 14" tires)）

Mass

Approx. 20 kg

Reducer type

Cycloid reducer

Reduction ratio

1/11

Motor type

Axial gap type permanent magnet
synchronization motor

Max. motor speed

15000rpm*New Product Development R&D Center New Product Development Dept.

-56-

Development of In-Wheel Type Integrated Motor Axle Unit

Counterweight

Wheel

Needle roller bearing

Axle unit

Tube forming

Hub bearing

Reducer

Motor

Fig. 1 Schematic of axle unit

Fig. 2 Installation of the unit

As shown in Fig. 4, the gears used are an externalgear that features an epitrochoidal curve gear tooth
profile and an internal gear with a circular arc tooth
profile (hereafter referred to as an outside pin). Only
autorotation motion is utilized with the inside pinsituated on the bore of the eccentrically oscillatingexternal gear.

3. Design concept(1) Application of reducer
Usually, the size of a motor is dependent on the
maximum torque generated. Since direct driving inwheel motor axle units do not have reducers, themotors must provide greater torque, making the
motors unavoidably larger. As a proportion of thewhole axle unit weight, motor weight is large.
Therefore, to design a compact, lightweight axle unit,
employment of a compact, lightweight motor isnecessary. We attempted to decrease the maximum
torque needed for the motor by incorporating a
reducer and selecting a high-speed motor so that asmaller motor could fulfill the requirements.

CBS(2) Type of reducerA

Fig. 3 Schematic of K-H-V type planetary gear

We sought a reducer type that helps satisfy thesmaller space requirement and provides a greater
reduction ratio. With a 2K-H planetary gear reducing
mechanism, which is a type often used as a reducer
system, the reduction ratio per stage is in the range of
1/4 to 1/6. To achieve a higher reduction ratio with thistype, at least two stages of gear rows are necessary,and this in turn leads to a need for a large,
complicated reducer.
In seeking a solution, we focused on cycloid
reducing mechanisms 1-3), all of which are K-H-V
planetary gear reducing mechanisms, such as the one
illustrated in Fig. 3. With this mechanism, the final
reduction ratio is dependent on the difference in thenumbers of teeth between the gears. Thus, even one
gear train can achieve a greater reduction ratio.

Outside pin(internal gear)

Inside pin(output shaft)External gear

Input shaft

Fig. 4 Basic structure of reducer-57-

NTN TECHNICAL REVIEW No.73（2005）

Use of this gearing arrangement allows many teeth
to remain in simultaneously meshed relations, and, asa result, transmission torque per unit volume can begreater, helping realize a compacter reducer design.
The hub bearing inner ring raceway groove isformed on the flange member to which the inside pins
are secured, and the flange member and the hub
bearing inner ring member with the hub flange are
fastened together using NTN's propriety tube forming
technique, which is also applied to our 4th generation
hub bearings. The tube forming technique has lead toa simple fastener structure between the hub and the
reducer, allowing further reduction of the axle unit
size.

with the same lubricant, the number of seals can be
decreased, simplifying the structure of the axle unit,
and the axial direction length of the axle unit can beshorter.
Though not shown in Fig. 1, a dog clutch or othertype of parking brake, a component absolutely
necessary for an axle unit, is located in the right endof the motor.

(3) Higher reducer efficiency(1) The efficiency of two reducer types was measured.
One reducer type was a rolling contact bearing
type that incorporated a needle roller bearing into
the sliding contact area between each inside pin
and outside pin set on the reducer. The otherreducer type was a sliding contact bearing typelacking incorporated needle roller bearings. As
shown in Fig. 5, at the input rotation speed of 3000
rpm, the rolling contact type reducer exhibited anefficiency approximately 5% better than that of thesliding contact type reducer (maximum efficiency of94%), showing the effectiveness of the rolling
contact bearing. This effect is nearly equivalent to
the effect obtained from a two-stage configurationof an involute gear type 2K-H planetary gear
reducer. An oil bath method was used for
lubrication.

4. Performance Test Results
The results of evaluation tests are described below
for the basic performance of a motor-less axle unitthat consists of a reducer and a hub bearing.

Generally, the transmission efficiency of a two-stage
2K-H planetary gear reducer with involute gears of areduction ratio of 1/10 to 1/20 is approximately 95%.
The transmission efficiency with a cycloid reducing
mechanism is said to be lower. One possible reason
for this is that when the eccentrically oscillatingexternal gear rotates, sliding contact occurs between
the outside pin and the external gear and between theinside pin and the external gear.
Therefore, we attempted to reduce the loss at themeshed portions by incorporating rolling contact
bearings into these sliding contact areas.

(4) Higher reducer speed
With a cycloid reducing mechanism, in principle, theexternal gear eccentrically oscillates and the inertial
force induced by this motion causes the vibration of
the component perpendicular to the rotation axis toincrease. Usually, therefore, another set of reducers(external gear) acting in reverse phase is incorporated
to dampen this vibration. However, the couple of force
resulting from unbalanced inertia on the axes of thetwo external gears perpendicular to the rotation axis
still remains, and the effect of the unbalance isprominent when the reducers run at higher speed. Our
axle unit incorporates a counterweight to dampen thevibration induced by this couple of force resulting from
unbalance inertia.

100

Efficiency of reducer [%]Input rotational speed=3000rpm(5) Bearing layout

90

80

Rolling contact typeSliding contact type70
0

The motor rotating shaft has a cantilever support
structure from the reducer housing. Though not shownin Fig. 1, the knuckle is formed around the reducer
outside periphery as a part of the reducer housing,allowing the wall thickness of the motor housing to bethinner.
Thanks to this bearing layout, the space from the
hub bearing to the motor bearing can be lubricated

2

4

6

Input torque [Nm]Fig. 5 Efficiency of reducer

-58-

8

10

Development of In-Wheel Type Integrated Motor Axle Unit

(2) Acceleration pickup devices were installed to theouter periphery and end face of the reducer
housing to measure the radial vibration and axial
vibration on the running reducer. The
measurement results are plotted in Fig. 6. Though
no load was applied to the reducer, it ran smoothly
without developing significant vibration at speeds
up to 15000 rpm.

5. Conclusion
Seeking to realize a compacter and lighter in-wheel
integrated motor axle unit, we adopted a cycloid
reducer mechanism. By replacing the sliding contactsof the reducer with a rolling contact arrangement, atransmission efficiency of approximately 94% wasachieved. We also verified that, though not under an
actual load, the reducer can run smoothly without
developing significant vibration at speeds up to 15000
rpm.
Since axle units are critical automotive components,
we are going to improve the reliability of our unit
through evaluation with various tests under actual
automobile operating conditions.

5

axial
radial

Acceleration [m/s2]

43

References

2(1) Muneharu Morozumi: Theory and design calculation
technique for planetary gears and differential gears,
Nikkan Kogyo Shinbun, Ltd., p1-6 (1989).
(2) Japan Institute of Plant Maintenance: Book of
reducers, Japan Institute of Plant Maintenance, p2226 (1994).
(3) Sumitomo Heavy Industries, Ltd.: Catalog C2001-5.

1

0
0

5000

10000

15000

Input rotational speed [rpm]Fig. 6 Vibration of reducer

Photos of authors

Minoru SUZUKI

Dawei WANG

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

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

-59-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Article ]Development of the Mono Ring CVT- Principle and Mesurement of Transmission Efficiency -Yuichi ITOH*
Tomoya SAKAGUCHI*

There has been several variations of CVT (Continuously Variable Transmission) designs over theyears. A CVT can give high efficiency, low fuel consumption and smooth drivability to automotive
applications by changing output ratios with no actual gear steps. However, the mechanism of a CVT
can be intricate and contain many parts.
We have designed a new CVT, which is compact, simple and has no additional outside controllers.
This CVT is composed of a ring with teeth, a set of V-pulleys, and an output cogwheel. It has been
named the Mono-Ring CVT (MR-CVT) because the single ring gear carries all the torque.
This paper reports on the configuration and principle of the MR-CVT. The efficiencies of a
prototype MR-CVT were also measured confirming the validity of the basic design.

This report introduces a compact CVT we are
currently developing that consists of a smaller numberof components and does not need an external control
system.
Though it resembles the Bayer CVT, our CVT canalso suppress spin at the contact points. In addition,
though similar to a belt type CVT because it employs
V-pulleys, our CVT uses them in the drive section only
and one toothed ring is inserted between the V-pulleys
to transmit the power to the output gear.
We call our CVT the Mono-Ring CVT (MR-CVT)
because a single (mono) toothed ring is used as themedium. The structure, operating principle andmeasured transmission efficiency of the MR-CVT aredescribed below.

1. Introduction
Various technologies have been invented forautomotive continuously variable transmissions(CVT).1) 2) Recently, multi-stage designs have been
increasingly used for automotive automatictransmissions (AT) in order to achieve bettertransmission of engine power and smooth gearchange. By adopting "stepless transmission," the mostadvanced form of multi-stage transmission, CVTs
realize higher efficiency, lower fuel consumption andextremely smooth drivability.

*New Product Development R&D Center Mechatronics Research Dept.

-60-

Development of the Mono Ring CVT

The V-pulleys are supported on the shafts of theballscrews by angular contact ball bearings. The pairof V-pulleys sandwich a toothed ring, where the ringhas with teeth around its circumference. The two
guide rollers direct the toothed ring and swing theoutput gear with it as a unit. This swing varies thecontact radius of the toothed ring with the V-pulleys,
making stepless speed change possible. The directionof rotation of the V-pulleys is regulated in the
counterclockwise direction only by means of a selfadjustment mechanism described later. The toothedring and the output shaft are linked with a spur gear.
The basic mechanism of the MR-CVT isschematically illustrated in Fig. 2. Since the input
torque is transmitted to the toothed ring by the V-

2. Mono-Ring CVT (MR-CVT)The internal structure of our prototype MR-CVT isillustrated in Fig. 1. The input shaft and the output
shaft are parallel to each other in a offset relation, andboth are supported by bearings fixed to the housing.
The input shaft is equipped with a pair of V-pulleys
that rotate concentrically and in synchronicity with the
input shaft, so the V-pulleys can shift in the axial
direction. The input shaft has ball splines to allow theV-pulleys to shift in the axial direction and transmit
torque smoothly. A right-hand and left-hand pair ofballscrews is included on both ends of the input shaft.
By turning the shafts of the ballscrews, the V-pulleys
can be shifted axially.

Guide roller
Decelerationside

V-pulleyg
Swin
Acceleration
side

Guide roller

Toothed roller

Ballscrew

Toothed roller

Angular contact
ball bearing

Guide roller

Ball spline
Input shaft

V-pulley

Output shaft
Output gear

Guide roller

Output gear

100mm

Fig. 1 Schematic view of MR-CVT

Gear pitch
V-pulleycircleGuide roller

Toothed roller

Guide roller
F1

Q

0g

2θ

0IN
Q

Rt
Rg

Toothed roller

Contact point between
toothed ring and V-pulleys

F3

V-pulley

Gear pitch circle
0OUT

Fig. 2 Basic mechanism of MR-CVT-61-

NTN TECHNICAL REVIEW No.73（2005）

pulleys, the toothed ring is subjected to the frictional
force F1.
The torque transmitted to the toothed ring is furthertransmitted to the output shaft gear through its teeth,
and, as a result, the reaction force F3 from thetransmission force acts on the toothed ring. The
relation between F1 and F3 can be described byformula (1), where Rt and Rg are the radii from thetoothed ring center at the F1 and F3 application points,
respectively. For accuracy, the pressure angles of thegears should be considered, but for convenience of
calculation, we took the pressure angles as zero.

Since the toothed ring is stable when the above
relationship is valid, a self-adjustment mechanism isformed in which the friction coefficient μis determinedfrom the slope angle θ of the V-pulleys even if
operating conditions vary.
This is because F1 and F3 vary according to the
input torque, so the contact force Q occurring between
the V-pulleys and the toothed ring automaticallyvaries.
As shown in Fig. 2, when Og-Oout and the Og-Oin
are perpendicular to each other, the optimal angle θof the V-pulleys can be determined based on thegeometrical relationship of the toothed ring alone andthe friction coefficient μ as described with formula (4).
However, as the toothed ring swings on the output
shaft, the Og-Oout and the Og-Oin are no longer
perpendicular as shown in Fig. 3. Since the contact
point between the toothed ring and the V-pulleys shifts
to a point on the prolonged line from the Og-Oin, therelation in formula (2) is no longer valid.
Fig. 3 shows a situation where the toothed gear isin the acceleration position (top gear side). Due to thefrictional force F1, component F4 that forces thetoothed ring into the V-pulleys occurs. Thus, theresultant force F5 is:Rg
F3=F1×―― ………………………………（1）Rt
The force F3 also functions to force the toothed ring
into the V-pulleys. The contact force Q caused by thisforce F3 and acting from the V-pulleys onto thetoothed ring can be described by formula (2) thattakes the slope angle of the V-pulleys as θ.
F3Q= ―――― …………………………………（2）2sinθ
If the friction coefficient at this contact area isassumed to be μ, then formula (3) holds.

F5=F3＋F4F1μ=―― ……………………………………（3）2Q

Fig. 4 illustrates the position of the toothed ring in
the deceleration position. The F5 in this situation is:

From formulas (1), (2) and (3), formula (4) is
derived.

F5=F3−F4
If angle θ between Og-Oout and Og-Oin is as shownin Fig. 3, then from this diagram, the contact force ψderived from the push-in force F5 can be described byformula (5).

Rtsinθ=μ×―― ……………………………（4）
Rg

F2

Rt

F2

F1
F1

Rg
F4
0in
F5

ψ

F4
F5

0g

0in

0g

F3

F3

0out

0out

Fig. 3 Speed-up condition

Fig. 4 Slow-down condition-62-

Development of the Mono Ring CVT

3. Test method

F5
F3＋F1 sinψF（R1t/Rg＋sinψ）
Q= ――――――― = ――――――― = ――――――――――
2sinθcosψ
2sinθcosψ
2sinθcosψTo verify the feasibility of the operating principle ofour MR-CVT, we developed a test rig and measured
the transmission efficiency of the MR-CVT.
The test rig is schematically illustrated in Fig. 6.
Both the input and output sides of the MR-CVT were
provided with a torque meter to measure the rotation
speed and torque. For performance measurement of
the MR-CVT, the transmission efficiency wasdetermined based on the workload, which is the
product of the rotation speed and torque of the input
and output power in normal operation. On the test rig,the drive motor and the load motor were linked with aninverter, and the energy regenerated on the load
motor was used for the drive motor.
The range of the gear ratio, the ratio of the rotation
speed of the input shaft to that of the output shaft, was
0.3 to 1.3. The gear tooth ratio of the toothed ring (88teeth) to the output gear (60 teeth) was 1.47. The test
conditions are summarized in Table 1.
The lubricant used was a market-available CVT oil,and the supply inlet temperature was adjusted to 50˚Cwith a temperature adjustment system.

…………………………………………………（5）If formula (3) is substituted into the formula (5), thenthe optimal V-pulley inclination θ relative to ψ can be
determined as a function of ψ by using formula (6)below.
μ（Rt/Rg＋sinψ）
θ= sin-1 ――――――――――cosψ［］ ………………（6）In other words, the MR-CVT can have optimal
efficiency if the slope angle θ, which is determined byψ, the friction coefficient μ and the geometry of the
toothed ring when it swings on the output shaft satisfy
formula (6). Incidentally, the relation involving ψ,
Oout-Og, Rt, Oout-Oin and the radius r from the centerof the V-pulleys to the contact point is determined by
the cosine theorem. In other words, the ψ can be
determined based on the CVT design data.
The relation between θ and r in our prototype MRCVT is illustrated in Fig. 5. As can be seen the graphas the radius of contact point r between the toothedring and the V-pulleys increases, the slope angle also
increases. Incidentally, we designed the subcurvatureof the toothed ring so that the contact surfacepressure between it and the V-pulleys stays below the
maximum allowable level when the input torque is at
maximum.

Slope angle θ, degrees

10

8

6
Torque meter

MR-CVT
Torque meter

4

2

Fig. 6 Test rig for measuring transmission efficiency
0

5

15

25

35

45

55

Table 1 Test condition

Radius of contact point r, mm

Input rotation speed, min-1

Fig. 5 Relation between radius of contact point andslope angle of V-pulley

Input torque, N-m
Transmission gear ratio (radius, mm)
Dynamic viscosity of lubricant
Lubricant supply temperature-63-

1000，2000，4000
10〜60
0.4〜1.2 (17〜50)
25.5 mm2/s @40˚C
5.12 mm2/s @100˚C
50˚C

NTN TECHNICAL REVIEW No.73（2005）

100

4. Test results

Gear ratio 0.4
Gear ratio 0.6

Transmission efficiency, %The measurement results for the input torque andtransmission efficiency in an input rotation speed
range of 1000 to 4000 min-1, at gear ratios of 0.4, 0.6,
0.8, 1.0 and 1.2 are illustrated in Figs. 7 through 9.
Transmission efficiency increased with a greater
input torque and with a higher gear ratio. At theordinary gear ratio for automobiles of 1.0 or greater,transmission efficiency of 85% or greater was attainedin almost all the torque range and the maximum
transmission efficiency reached 92%. Though thetransmission efficiency at the gear ratio of 0.4 waslow, this gear ratio is used only when an automobile
starts to run. The time when this gear ratio is usedaccounts for a smaller portion of the whole duration of
car usage, so this low efficiency does not seem togreatly affect the overall efficiency of this MR-CVT in
actual automobiles.
The reason for the increased transmission efficiency
with higher gear ratios seems to be a decrease in theloss that results from spin-induced slip in the contact
area between the toothed ring and the V-pulleys. The
loss of input torque at 10 N-m seems to result from agreater ratio of various frictional losses to transmission
efficiency in the MR-CVT.

95

Gear ratio 0.8
Gear ratio 1.0
Gear ratio 1.2

90

85

80

75

70
0

10

20

30

40

50

60

Input torque, N・mFig. 7 Input torque and transmission efficiency(1000min-1)
100
Gear ratio 0.4

Transmission efficiency, %Gear ratio 0.695Gear ratio 0.8
Gear ratio 1.0
Gear ratio 1.2

90

85

80

75

70
0

20

40

60

Input torque, N・mFig. 8 Input torque and transmission efficiency(2000min-1)
100
Gear ratio 0.4

Transmission efficiency, %Gear ratio 0.695Gear ratio 0.8
Gear ratio 1.0
Gear ratio 1.2

90

85

80

75

70
0

20

40

60

Input torque, N・mFig. 9 Input torque and transmission efficiency(4000min-1)

-64-

70

Development of the Mono Ring CVT

5. Conclusion
We have invented a simple, compact MR-CVT on
which a single (mono) toothed ring swings to changetransmission speed. The MR-CVT features a selfadjustment mechanism that is the result of the designof the V-pulleys, guide rollers and other components.
We made the following findings in the tests detailed inthis report.
(1) Our MR-CVT functions correctly without needfor an external pressure mechanism due to its
self-contained adjustment capability.
(2) The transmission efficiency of our MR-CVT wasat least 85% with a gear ratio of 1.0 or higher
over nearly the entire speed range and the
maximum transmission efficiency reached
92%.

References
1) S.H.Loewenthal：Advanced Power Transmission
Technology NASA cp-2210（1981）79.
2) The Japan Society of Mechanical Engineers: P-SC62
Work Group for Investigation into Traction Drive(1985).

Photo of authors

Yuichi ITOH

Tomoya SAKAGUCHINew Product Development
R&D Center
Mechatronics Research Dept.

New Product Development
R&D Center
Mechatronics Research Dept.

-65-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Hybrid Grease NA204F
for Automotive Electrical Instruments and Auxiliary Device

Masaki EGAMI***
Mitsunari ASAO***
Tomoaki GOTO***This report introduces newly developed Hybrid Grease NA204F for automotive electricalinstruments and auxiliary devices. NA204F has a unique composition that is a mixture of urea
grease and fluorinated grease, and shows longer grease life at high operating temperature (up to
200˚C) at lower cost than that of typical fluorinated grease. Moreover, NA204 is greatly improved on
weak points of fluorinated grease, including enhanced rust prevention ability and greater resistanceto grease leakage from the bearing. NA204F is expected to be applied to several kinds ofautomotive applications.

auxiliary devices.

1. Introduction
To cope with the recent trend toward higher
performance and increased power in automobiles, theneed has been mounting for automotive electricalinstruments and auxiliary devices that are morecompact, lighter, faster and have greater functionality.
As a result, the operating conditions for automotive
bearings have become increasingly severe. The
greases for bearings on automotive electricalinstruments and auxiliary devices often consist of urea
compounds as thickeners and synthetic oils as baseoils. The greases used for applications in ambient
temperatures of 180˚C or higher sometimes use
fluorine compounds as thickeners and base oils.
Greases containing fluorine compounds are much
more expensive than urea greases. For this reason,strong demand exists for a less expensive, hightemperature long-life grease.
To address this problem, NTN has developed
Hybrid Grease, which features excellent costperformance and high functionality. This paper
describes the features of the Hybrid Grease and basic
characteristics associated with its application inbearings for automotive electrical instruments and2. Hybrid Grease defined
Grease prices are proportional to high-temperature
life values as summarized in Fig. 1. In other words,
greases that are more expensive also boast longer
lives. With the goal of ending this commonly accepted
relationship, we have been committed to developing aless expensive, high-temperature, long life grease for
automotive electrical instruments and auxiliary7000

180˚C grease life, h

6000
5000

Fluorine grease

4000

Lithium soap/ester
3000

Urea/aromatic ester

2000

Urea/ether

1000

Urea/PAO, urea/ester

0
0

5000

10000

15000

20000

25000

30000

Price, yen/kg

Fig. 1 Cost and high temperature life of greases**New Product Development R&D Center Mechatronics Research Dept.
**Automotive Sales Headquarters Automotive Engineering Dept.

-66-

Hybrid Grease NA204F

devices. As a result, we have succeeded indeveloping our Hybrid Grease.
Our Hybrid Grease is essentially a mixture ofdifferent greases. The one introduced in this paper isa mixture of urea grease (thickener: urea compound,base oil: ester oil) and fluorine grease (thickener:
ethylene tetrafluoride resin powder (PTFE), base oil:perfluoropolyether oil (PFPE)). Generally, mixing
dissimilar greases needs to be avoided since it often
leads to changes in grease properties such as
consistency and reduced grease life. 1) However, ourattempt to mix urea and fluorine greases was an
exception because we have attained extremely
favorable grease properties.
Fig. 2 illustrates the trend in grease life with various
mixture ratios of urea and fluorine greases. The test
conditions were in accordance with ASTM D 3336. A
6204 bearing with a shield plate, whose internal space
was 38% filled with grease, was loaded onto thegrease life test rig illustrated in Fig. 3, the test rig wascontinuously run at 200˚C, to evaluate the time until
bearing seizure. As can be seen in Fig. 2, the life of
urea grease alone is extremely short. However, thelife of a mixture of urea grease and fluorine grease islonger than that of fluorine grease alone. This meansthat high-temperature life equivalent to or better thanthat of fluorine grease can be attained at a lower price.
From various analytical results, we estimated thelong life mechanism of the Hybrid Grease to be as
summarized in Fig. 4. At 200˚C, urea grease
deteriorates in a short time span, and reaches the endof its life in approximately 200 hours.

Grease life, h（200˚C）

6000
5000
4000
3000
2000
1000

0
0
(Urea 100%)

20

40

60

80

100
(Fluorine 100%)

Fluorine grease compounding ratio, VOL %Fig. 2 High temperature life of Hybrid Grease

Test bearing: 6204 (w/ shield plate)Amount applied : 38% of internal space
Temperature: 200°C (bearing outer ring end face)
Inner ring speed : 10000 min-1
Load: Fa=Fr=67 N
Cartridge heater
Test bearing
Loading coil spring
Support bearing
Thermocouple

PulleyFig. 3 Grease life test rig

Cage

Outer ring
Shield plate

Urea grease

Inner ring
AppliedgreaseRolling elements (balls)
Start of operation→
Deterioration and evaporationof urea grease→
Seizure and end of life

Before operation

Fluorine grease

Corrosion-induced wear

※Grease channeled tothe sides does notcontribute to lubrication.
Progress of grease deteriorationin rolling area→Reaction with rolling
elements and rolling surfaces→
Progress of wear

Start of operation→
Channeling

HybridGrease※Residual grease
principally consists offluorine grease.

Start of operation→
Deterioration of urea
greaseSeizure andend of life

※Full utilization offluorine grease
Film formation is a self-repairing feature

Film formation on rolling elements androlling surfaces produced by deteriorationof urea grease→Inhibition of reaction
between fluorine and steel surface→
Grease deterioration and wear are inhibited

Fig. 4 Estimated mechanism for long grease life of Hybrid Grease-67-

NTN TECHNICAL REVIEW No.73（2005）

Immediately after the start of bearing rotation, mostof the fluorine grease is shifted to the sides of the
raceway, and only a small amount of grease remains
on the rolling surfaces (channeling 2)). From theanalysis of the bearings that underwent the test, wefound that the grease adhering to the inner side of theshield plate does not contribute to lubrication and therolling surfaces are lubricated with the small remaining
amount of grease. When the grease on the rollingsurfaces deteriorates and the supply of base oil to therolling surfaces and cage pockets is insufficient, wear
(corrosion-induced wear) of the steel members
progresses due to the chemical reaction between thesteel surfaces and PFPE 3) 4) 5), and the bearing
approaches seizure and the end of life.
However, as the urea grease blended in our Hybrid
Grease degrades, a film is formed on the rolling
elements and rolling surfaces that inhibits the reaction
between the steel parts and PFPE, preventing wearon them. Thus, the heat resistance of PFPE is fully
developed to help extend the high-temperature life of
the Hybrid Grease.

3.1 General properties and high-temperature life
The general properties of our newly developed
grease are summarized in Table 1. For comparison,
this table also provides the properties of a hightemperature, long life fluorine grease A that has been
performing well for automotive electrical instruments
and auxiliary devices and of fluorine grease B, whichfeatures excellent cold temperature performance.
Compared with fluorine grease A, Hybrid Grease
NA204F features a high-temperature life
approximately 1.5 times as long, as well as a lower
low-temperature torque. Compared with fluorine
grease B, our Hybrid Grease boasts a hightemperature life approximately six times as long,though its low temperature torque is somewhat
inferior.

3.2 Cold temperature noise characteristics
Depending on the pulley specifications andoperating conditions, a belt pulley incorporating a
deep-groove ball bearing can develop cold
temperature noise known as hoot noise when started
in a cold climate. With a frequency range of 500 to1000 Hz and a sound pressure that can exceed 100
dBA, this noise is very offensive.
The cause of hoot noise seems to be that the pulleysystem develops resonance due to self-excited
vibration of the rolling elements, causing the outer ringto vibrate in the axial direction (translational motion).
Some types of grease are known to greatly promotenoise generation.
For evaluation (see test conditions in Table 2), thetest bearing incorporated into the test pulley was cooledin advance for 5 hours in a low temperature bath kept at-40˚C, then the pulley was installed in the test rig
illustrated in Fig. 5 that was at a room temperature.

3. Application to automotive electricalinstruments and auxiliary devices
Hybrid Grease compositions can have various
fluorine grease mixture ratios as shown in Fig. 2. We
determined the mixture ratio that leads to the longest
grease life. Beginning with grease with this ratio as thebase, we blended optimal amounts of additives toimprove characteristics including low temperature
performance and rust preventive performance inaddition to high-temperature life, resulting in thedevelopment of Hybrid Grease NA204F for
automotive electrical instruments and auxiliary devices.

Table 1 General properties of Hybrid Grease
Unit

Hybrid Grease
NA204F

Thickener−

Diurea compound
PTFE

PTFE

PTFE

−

Base oil−Special ester oil
PFPE
PAO

PFPE

PFPE

−

Characteristics

Fluorine grease A Fluorine grease B

Test method

mm2/s

148

160

90

JIS2220.23

60-times worked penetration−

Category 1（310〜340）

Category 2（265〜295）

Category 2（265〜295）

JIS2220.7

Dropping point

˚C

more than 250−−

JIS2220.8

mass %

0.0

4.8

4.7

JIS2220.11

1.22

1.85

1.94

−

Base oil dynamic viscosity (@40°C)

Oil separation (100˚C×24 h)Specific gravity
Low temperature torque Starting(-30˚C) SteadyGrease life＊

N・m
h

Start

16.9

37.6

4.9

Running 7.0

28.0

2.6

6000

4050

965

＊Evaluation test shown in Fig. 3-68-

JIS2220.18
Per ASTM D3336

Hybrid Grease NA204F

Belt tension was applied and operation was started
when the bearing temperature rose to -20˚C and wasmaintained until the bearing temperature reached
10˚C. During this period, the noise level (coldtemperature noise) was checked by listening. For
each grease, evaluation was repeated 5 times foreach of two bearings (n=10).
As can be seen by the data in Table 3, like thefluorine grease B that excels in cold temperature
performance, our Hybrid Grease NA204F did not
develop cold temperature noise and was found to
exhibit excellent cold temperature noise prevention
ability.

3.3 Rust prevention ability
Since ordinary rust preventing grease additives do
not dissolve in PFPE, they cannot be used in fluorine
greases. As a result, the rust prevention ability offluorine grease is inferior to that of other grease types.
Since a rust-preventing additive is dissolved in theurea grease base, however, our Hybrid Grease has
powerful rust prevention ability.
Fig. 6 summarizes rust occurrence percentages
obtained from an evaluation test conducted according
to ASTM D1743 in which salt water was used insteadof pure water. For evaluation, a 4T-30204 tapered
roller bearing coated with each of the test greases
was immersed in 1% aqueous sodium chloride
solution for 10 seconds, and was allowed to stand for
48 hours in an atmosphere of 40｡C and 95% relative
humidity. Then, the rust spots on the outer ring rollingsurface were counted. The rolling surface wascircumferentially divided into 32 equal segments, andthe number of segments with rust was divided by 32 todetermine the rust occurrence rate.
Compared with the rust prevention ability of fluorine
greases A and B, Hybrid Grease NA204F exhibited
excellent rust prevention ability.

Idler pulley
Test pulley

Driving pulley
Timing belt
60

Rust occurrence %Fig. 5 Tension pulley type test rig

Table 2 Test condition of hoot noise at low temperature
6203 (light-contact type, w/ rubber seal)

Test bearing
Radial clearance, μm

12〜16Amount of applied grease

38% of internal space-1

Outer ring speed, min2700

Pulley load, N

127−20〜10

Number of repetitions

n=2×5 times

Hybrid Grease
NA204F
Fluorine grease A

Fluorine grease B

30
20
10

HybridGreaseNA204F

Number of repetitions Occurrence
1

2

3

4

5

No.1

○

○

○

○

○

No.2

○

○

○

○

○

No.1

○

○

×

×

×

No.2

○

△

×

×

×

No.1

○

○

○

○

○

No.2

○

○

○

○

○

Fluorine grease
A

Fluorine grease
B

Fig. 6 Rust prevention ability of greases

Table 3 Rate of hoot noise generation at low temperature
Item

40

0

Test temperature, ˚CGrease

50

％
0

60

0

Symbols〇: No noise occurrence△: Noise other than hoot noise (grease sound)
×: Hoot noise-69-

NTN TECHNICAL REVIEW No.73（2005）

3.4 Leakage characteristics of greases

4. Compatibility with seal rubber material

Fluorine grease is not compatible with the base oil(mineral oil) in rust-preventive oil that is used in the
bearing manufacturing process. As a result, if fluorine
grease is inserted in a bearing coated with such rustpreventive oil, the fluorine grease will slip on the oilfilm and may leak out of the bearing in a short periodof time. To prevent this problem with fluorine grease
application in bearings, it is common to limit theamount of residual rust preventive oil within the
bearing.
By following a manufacturing practice for applying
common grease, we inserted the same amount ofgrease into a bearing whose inside was coated with arust-preventive oil (internally rust-proofed bearing) andinto a degreased bearing. Each bearing was run for 10
minutes. The resultant grease leakage percentages(expression 1) are summarized in Fig. 7.

The rubber seal material in bearings with fluorine
grease is usually a fluorine-containing rubber in orderto satisfy heat-resistance requirements. The mostcommonly used fluorine-containing rubber materials
are so-called FKMs (an ASTMD1418 abbreviation
representing fluorine-containing rubber) that include abinary copolymer of vinylidene fluoride andhexafluoropropylene (VDF-HFP) and a ternary
copolymer comprising VDF-HFP andtetrafluoroethylene (VDF-HFP-TFE). These common
FKMs exhibit sufficient durability when combined withfluorine grease and do not pose any problem.
However, when an FKM contacts urea grease in ahigh-temperature situation, its vinylidene fluoride
component is attacked by the urea thickener andamine, developing a cross-linking reaction. As aresult, the FKM becomes hardened, its strength isdegraded and its sealing performance is eventually
lost. Our Hybrid Grease also affects common FKMs inthe same way as urea grease.
To address this problem, the authors developed anovel fluorine rubber seal material, NUF (NTN antiurea fluorine rubber), with a base rubber that consistsof fluorine rubber made of components immune tourea. Fig. 8 illustrates the time-dependent variationsin hardness and tensile strength with dumbbell test
pieces made of NUF and common FKM that were
immersed in Hybrid Grease NA204F at 200˚C. While
the physical properties of the common FKM greatly
degraded in a short time, those of the NUF did notdegrade much, remaining stable for a prolonged time.
This feature of FKM is valid not only when used with
Hybrid Grease, but also when it is used together withurea grease.

amount of leakage
Leakage percentage＝―――――――――――――×100 (%)
original inserted amount
Expression (1)Large amounts of fluorine greases A and B leaked
from the internally rustproofed bearing. In contrast,
only a small amount of Hybrid Grease NA204F
leaked from the internally rustproofed bearing.
As mentioned in Sec. 3.3, with its powerful rustprevention ability, Hybrid Grease will help contribute tobearings that are more resistant to rust compared tofluorine grease-filled bearings.

Grease leakage %1210Degreased bearing
Internally rust-proofed bearing

8

5. Conclusion

6
4

By combining urea and fluorine greases, we havesuccessfully developed Hybrid Grease NA204F, whichfeatures exceptionally long high-temperature life thathas no precedent in conventional grease technology.
At a temperature as high as 200˚C, its life is longer
than that of conventional fluorine greases. At thesame time, it is free from the drawbacks associatedwith fluorine greases, including poor rust prevention
ability and a tendency toward leakage induced by rustpreventive oil within bearings. Furthermore, it helps
prevent occurrence of cold temperature noise whenapplied to pulley bearings. Being less expensive than
fluorine greases, Hybrid Grease NA204F boasts
extremely high cost performance, and when used
together with our newly developed urea-resistant20HybridGreaseNA204F

Fluorine grease
A

Fluorine grease
B

Fig. 7 Leak rates of greases during rotation

Test bearing: 6204 (w/ non-contact rubber seal)
Applied amount : 38% of internal space
Inner ring speed : 10000 min-1
Load: Fa=39 N
Temperature: room temperature
Operating time : 10 min.

-70-Hybrid Grease NA204F

ning
Harde face
r
u
s
f
o
n
matio
Defor

20

Variation, %15

NUF
Common
FKM

10
5
0

100

200

300

400

500

600

700

800

900

1000

0
-5
-10

200˚C, in NA204F

Immersion time, h(a) Hardness
0

100

200

300

400

500

600

700

800

900

1000

0

Variation, %

-20
-40

ning
Harde faceof sur ation
m
Defor

-60

NUF
Common
FKM

-80
-100

Immersion time, h

200˚C, in NA204F(b) Tensile Strength

Fig. 8 Deterioration of fluorinated rubbers dipped into Hybrid Grease at 200˚C2) Japanese Society of Tribologists: Tribology
Handbook, Yokendo Co., Ltd. (2001) 715.
3) Shigeyuki Mori, Mariko Ito, Yuji Shitara, Yuji Enomoto:
Relation between lubricity and molecular structurewith perfluoroether oil (2nd report), Tribologist, 40, 15(1995) 923.
4) William R. Jones Jr., Mariko Ito (translation),
Shigeyuki Mori (translation): Perfluoroether-based
lubricants, Tribologist, 42, 3 (1997) 205.
5) Mineo Suzuki, Philippe Plat, Koji Matsumoto: JAXA
research and development report, Performance ofball bearings having a retainer impregnated withspace lubricants and their outgassing properties,
JAXA-RR-03-014 (2004).

fluorine rubber, NUF, it does not develop
incompatibility problems with seal rubber material.
Thus, we believe that Hybrid Grease NA204F will
greatly contribute to performance enhancement andcost reduction for automotive electrical instruments
and auxiliary devices.

References
1) Tadashi Hatakeyama: Mixing of dissimilar greases,
and variation in properties, Junkatsu, 19, 4 (1974)
331.

Photos of authors

Masaki EGAMI

Mitsunari ASAO

Tomoaki GOTO

New Product Development
R&D Center
Mechatronics Research Dept.

New Product Development
R&D Center
Mechatronics Research Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

-71-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Ballscrew for Automated Manual Transmission

Koji TATEISHI**
Keisuke KAZUNO**

Recently , the use of automated manual transmissions (AMT) are expanding in the automotive
market. Of the many methods of automation in the market, the ball screw and driven by DC motor
are methods that have the quickest response and largest system output.
Mitsubishi FUSO truck & bus and Bosch Co. ltd have developed a new small and medium sizedtruck that incorporates the AMT system. NTN has started to supply ball screw and support bearings
for this application.
This article documents the development and future applications.

1. Introduction
automated manual transmission (INOMAT-@) in
which the actuator is driven with ball screws. Thisactuator uses NTN ball screws and support bearings.
This paper describes the ball screws specificallydeveloped for this AMT.

Recently, an increasing number of automobiles
models have been incorporating automated manualtransmissions (AMT) to improve fuel efficiency andride comfort.
In the past, most automated gear change systems
for AMT were driven hydraulically or pneumatically.
These systems have recently been superseded byelectric motors in order to achieve more compact,
lighter weight units. Various attempts had been madeby combining electric motors with reducers that are
centered on worm gears or sliding screws. theseattempts had drawbacks including higher cost due tothe increased number of components, inability to offer
greater driving power due to relatively large friction,
lower efficiency, and poor response.
To solve these drawbacks, Mitsubishi Fuso Truck
and Bus Corporation and Bosch Corporation
developed a gearshift actuator for a compact truck

2. Structure of automated manualtransmission
2.1 AMT appearance
Photo 1 shows the AMT appearance. The gearshift
actuator is mounted to the top of the transmission,
which had previously required manual operation forgear shifting. This automatization has been achieved
without significant change to the transmission
structure. The significantly altered area is shownencircled by a dashed line.

*Automotive Sales Headquarters Automotive Engineering Dept.

-72-

Ballscrew for Automated Manual Transmission

Transmission
body

Gear shift
actuator

3. Features of ATM ball screws
The ball screws used in this ATM are shown in
Photo 2.

For shift side

Photo 1 Transmission
For select side

2.2 Overall layout of gearshift actuator
Fig. 1 illustrates the general structure of thegearshift actuator.1)
The gearshift actuator achieves gear change bysynchronously driving the two ball screws
perpendicular with each other by means of DC
motors. The gearshift actuator is capable of rapid
response thanks to the advantages of low friction ballscrews. The response times are 0.06 sec. for the shiftside, and 0.08 sec for the select side.

Sliding lever

Shifting shaft

Photo 2 Ball screw for AMT

3.1 Ball screw circulation sections
The circulation section on each of the ball screws is
a deflector type. Thanks to a new staking method
each deflector Photo 3. is very compact with a radial
dimension approximately 20% smaller than ordinary
return tubes.
The deflector staking assembly is shown in Photo
4. The deflector is inserted from the inside of the nut,and the wing is fit into the thread groove where theballs do not roll, and then the wing is fixed by staking(Photo 5). This structure is less expensive comparedwith conventional gluing or assembly with a screw, but
is reliable because it has sufficient strength against
loosening of the deflector member toward the outsideof the nut. The deflector is formed with high-precision
metal injection molding (MIM), guaranteeing stableball circulation.

Ball screw

Select motor

Non-contact
sensorBall screw

Shift motor

Non-contact
sensorFig. 1 Construction of gear shift actuator

Wing

Photo 3New deflector

Photo 4
Appearance of deflector
staking assembly-73-

Photo 5
Appearance of deflector staking

NTN TECHNICAL REVIEW No.73（2005）

Table 1 Specification of ballscrew

3.2 Durability
When the gearshift actuator is running, the forkshaped arm of the shift axle and the trunnion on theball nut come into contact and form a point ofapplication (Fig. 2).
If the arm and the nut trunnion are in an idealcontact relationship, the ball screw receives a pureaxial load. However, in reality, a moment load acts on
the ball screw from the trunnion because of factors
including the gearshift unit clearance and theinclination induced by the ball screw backlash.
As a result, greater load occurs on the balls situatedat both sides of the nut. Therefore, to ensure a longer
ball screw life, improving the load distribution on theballs is important.
To address this problem the authors have changed
the ball circulation system from a return tube system
to a deflector system that optimizes the span andcircumferential phase of the circulation deflector
relative to the position of the trunnion.
Comparison of both ball circulation systems in termsof ball contact pressure has been examined under theconditions summarized in Table 1.

Fork-shapedarm

Tube type

Deflector type
φ14.5

Lead (mm)4.0Ball dia. (mm)φ2.778
3.5×1

Number of circulations (rounds×rows)
Moment load (N-m)

1×4
17

The results of the test are summarized in Fig. 3.
There are two main advantages to application of adeflector system. 1) The deflector can be arranged in alonger nut-axis direction, and 2) the maximum numberof balls can be placed in the phase where momentload is applied. As a result, the contact pressure on
balls can be reduced by as much as 20%. Sufficientdurability was not attained with any return tube
system, but the NTN deflector system has achieved
the necessary durability.

4. Evaluation test
We developed a unique purpose-specific test rig
that can simulate the possible operation conditions ofactual automotive AMTs, and evaluated our ballscrews with it.

Fork-shapedarm

Shift shaft

4.1 Operability (efficiency)

Nut trunnion

Shift side ball screws must be capable ofdeveloping sufficiently large thrust because they aremeshed with a shift gear. Fig. 4 summarizes theresults of the relationship between the ball screw
backlash and the resultant thrust.

Fig. 2 Driving way of ball nut

Contact pressure reduction measure◆ Tube type▲ Deflector type

5000

Contact pressure (MPa)

Shift shaft

Characteristics
Shaft dia. (mm)

16%

20%
4000

3000

2000

1000

0
-15

-10

-5

0

5

10

15

Coordinate in ball axis direction (mm)※The loading position is taken as zero.

Fig. 3 Comparison of ball contact pressure-74-

Table 3 Durability test condition of select side ballscrew

2000
1900

Test

1800
1700
1600
1500
y=1191.9x+1536.2

1400
1300
1200
1100
1000
0.06

Durability test

Thrust on ball screw N

Ballscrew for Automated Manual Transmission

0.08

0.10

0.12

0.14

0.16

0.18

0.20

Radial backlash on ball screw mm
Fretting test

Fig. 4 Relationship between ballscrew's lash and thrust

Because a smaller radial backlash on any ballscrew reduces the degree of freedom of the nut, thefriction on the sliding surface of the actuator will
increase and as a result the thrust will decrease.
When the backlash on the ball screw is large, thelife of the actuator can become shorter because of the
moment load.
From the test results, we determined an optimal
backlash on the ball screw that ensures thrustnecessary for shifting action.

Durability test

Stroke±7mm

Drive motor

24 VDC motor

Ambient temperature

110˚C

Lubrication

Contaminated transmission oil

Number of operations

2,200,000

Vibration frequency

25〜400Hz

Acceleration

3.7〜96.1m/s2

Ambient temperature

Room temperature

Lubrication

Contaminated transmission oil*In three directions

In this report, we have introduced unique ballscrews that are used for novel, compact, lessexpensive actuators that are used in mechanicalmanual transmissions for compact and midsizedtrucks. We have developed these ball screws throughdesign improvements and innovations for the ball
circulation section. By applying these improvements,
NTN is currently developing ball screw units for use in
the electrically driven automotive actuators of engine
brake controllers in an attempt to expand the scope of
applications for these ball screws.

Reference

Table 2 Durability test condition of shift side ballscrew

Strength test

Inertia load by acceleration and deceleration of nut

5. Conclusion

The tests and test conditions for evaluating the
durability of the ball screws are summarized in Tables
2 and 3. When the shift gears start meshing, on rareoccasions an impact load may act on the shift side ballscrew, so a strength tests was performed on this ballscrew. The ball screw successfully passed the
durability test.

Fretting test

Max. load

Duration of vibration 12h

4.2 Durability1) Nakamura, et al. : Development of new INOMAT,
Mitsubishi FUSO Technical Review (2004, No. 1), p.
49-54

Test conditions

Test

Test conditions

Max. load

2800N

Stroke±5mm

Drive motor

24V DC motor

Ambient temperature

110˚C

Lubrication (grease) Ether-based synthetic oil + aromatic diurea
Number of operations

2,310,000

Impact load

8000N (two directions)Ambient temperature

Room temperature

Photos of authors

Lubrication (grease) Ether-based synthetic oil + aromatic diurea
Number of operations

2000

Vibration frequency

25〜400Hz

Acceleration

13.7〜96.1m/s2

Ambient temperature

Room temperature*in three directions: up-down,right-left, front-rear

Lubrication (grease) Ether-based synthetic oil + aromatic diurea
Duration of vibration 12h-75-

Koji TATEISHI

Keisuke KAZUNO

Automotive Sales Headquarters
Automotive Engineering Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Drawn Cup Needle Roller Bearings for Throttle Bodies

Hideki AKAMATSU**

Drawn cup needle roller bearings have an outer ring that is precisely drawn from a thin steel plate.
Of all bearings with outer rings, drawn cup needle roller bearings have the smallest cross sectionalheight, which enables both space and cost savings.
This chapter introduces the features as well as the performance of sealed, low friction, low air leakdrawn cup needle roller bearings for use in throttle bodies.

1. Introduction

2. Structure and features

A throttle valve is a device that controls the amountof air fed into an engine. To cope with recent growing
needs for better fuel efficiency and lower emissions,electronically controlled throttle valves that arecapable of fine adjustment of an air/fuel ratio are being
used increasingly.
As a result of the employment of electronic control,
the means for driving throttle valves has changed fromwires to motors. Rolling contact bearings, whichfeature decreased friction, are also being used
increasingly to enable both a compact lightweight
design and an improved response with the motor used
for this purpose.
Rolling contact bearings are being used in thesystems that control the airflow into automotiveengines as well. The support for the throttle blade
must not allow air to leak, therefore rolling contact
bearings must have good sealing performance.
To address these needs, we have developed drawn
cup needle roller bearings for throttle bodies that haveseals with lower friction and less air leakage. This
paper introduces the features and performance of
these drawn cup needle roller bearings.

The rolling bearings used for electronically controlled
throttle valves can be categorized as deep grooved ballbearings (hereafter, referred to as "ball bearings") anddrawn cup needle roller bearings (hereafter, referred
to as "drawn cup bearings") (Fig. 1).
Compared with ball bearings, drawn cup bearings
have smaller outside diameters, and as a result, theradial space of the housing can be smaller and the
housing can be lighter. Ball bearings are narrower
than drawn cup bearings, but their built-in seals alone
cannot completely prevent air leakage so additional
seals must be incorporated, which leads to the needfor additional space and expense.
Our newly developed drawn cup bearing assembly
consists of a drawn cup bearing fabricated by the prebent technique* and a seal that does not have a core
metal. This bearing does not need additional sealing,
and it can contribute to environmental protection
through cost reduction, space conservation andreduced weight.
*Pre-bent technique: a technique in which a cage and rollers are

fit into a drawn cup outer ring, and then the assembly issubjected to heat treatment.

*Automotive Sales Headquarters Needle Roller Bearings Engineering Dept.

-76-

Drawn Cup Needle Roller Bearings for Throttle Bodies

Throttle blade

Throttle blade

Return spring

Motor

Return spring
Separate seal

Motor
Drawn cup bearing(w/ built-in seal)

Deep grooveball bearing

Control gearing

Drawn cup bearing application example

Control gearing

Deep groove ball bearing application example

Fig. 1 Structure of electronic throttle bodies

Table 1 Specification of drawn cup needle roller bearings
for throttle bodies

Development
targetSpecifications

Dimensions Inscribed circle dia. 10×outside dia. 14×width 14 mm

Fig. 2 Structure of drawn cup needle roller bearings
for throttle bodies

Outer ring

Chromium molybdenum steel

Cage

V-type welded cage

Rollers

Crowned

Seal

Fluorine rubber, no core metal

Rated load

Dynamic rated load Cr 4500 N
Static rated load
C0r 5100 N

OthersHigh performance grease applied

Friction

1.1 N-cm max.

Air leakage 0.8cm3/min. max.

3. Friction Performance
Table 2 Test result of friction

The friction performance of the bearings that
support the throttle blade is a critical factor in adopting
a compact design and achieving quick response withthe motor.
A friction measurement value is obtained byinserting a torque gauge with a shaft of a specified
diameter into a drawn cup bearing that is press-fitted
into a standard ring, and turning the gauge 90
degrees. The friction value of these bearings isapproximately 0.8 N-cm, which is sufficiently lower
than the development target of 1.1 N-cm (Table 2)-77-

Development target

1.1 N-cm max.

Number of samples (n)30 pcs.

Mean value

0.78 N-cm

Max. value

0.85 N-cm

Min. value

0.70 N-cm

Standard deviation

0.054 N-cm

NTN TECHNICAL REVIEW No.73（2005）

4. Air Leakage Prevention Performance
Because the throttle valve regulates the amount of
air fed into the engine, the amount of air leakage fromthe supports of the throttle blade must be strictly
controlled.
The measuring method and resultant
measurements of the air leakage on drawn cup
bearings are summarized below.
Before starting the air leakage measurements, eachdrawn cup bearing was press-fit into a ring of specified
dimensions, and the shaft was inserted into thebearing of a specified diameter. This set was thenmounted to a sufficiently rigid measuring head. The
other measuring head was shut with a lid, and wasused as a standard control. These heads wereconnected to an air leakage tester, the open space ineach head was adjusted to a negative pressure, andthen air leakage was measured. Photo 1 shows the
air leakage tester, and Fig. 3 illustrates themeasurement heads.
The amount of air leakage with our drawn cup
bearing was 0.01 cm3/min. or lower, which is
sufficiently lower than the development target of 0.8
cm3/min (Table 3).

SuctionStandard control

Negativepressure
Test bearing
SuctionAtmosphericpressure

Fig. 3 Measurement part
Table 3 Test result of air leakage
Bearing Diameter of shaft
Air leakage
inserted (mm)No.
Development target: 0.8 cm3/min. max.
0.01cm3/min. max.

No.1
No.2

φ9.9750.01cm3/min. max.

No.3

0.01cm3/min. max.

No.4

0.01cm3/min. max.

5. Conclusion
This paper has introduced our unique drawn cup
bearing assembly for a throttle valve. This assembly
incorporates a compact, less expensive drawn cup
bearing and a seal that features lower friction and air
leakage prevention capabilities. Use of this bearing
helps reduce the size of the throttle valve and improve
the response speed. We believe that this bearing willoffer better control of air-fuel ratios for automotiveengines, helping realize better fuel efficiency andlower emissions and thus contributing to global
environmental protection.
Photo 1 Air leakage tester

Photos of author

Hideki AKAMATSU
Automotive Sales Headquraters
Needle Roller Bearings
Engineering Dept.

-78-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Compact and Shudderless Constant Velocity Joint (EPTJ)

Tatsuro SUGIYAMA**
Yuuichi ASANO**NTN Corporation has successfully developed a new, high efficiency Pillow Journal Tripod Joint(EPTJ). This type of joint drastically improves the NVH (Noise, Vibration, Harshness) characteristic
in vehicles. The high efficiency E-series PTJ, what we call EPTJ, is also lighter and more compact
than a conventional PTJ. The EPTJ maintains the excellent vibration characteristics of the ultra-low
vibration plunging tripod joint PTJ that NTN brought into the market in 2002. Moreover, it is smaller(4% reduction in outer diameter) and lighter (weight reduced by 8%) than the PTJ.

result, the induced thrust from the running PTJ has
been reduced by more than 70% (at an operating
angle of 6˚ or greater) compared to conventional
tripod joints, greatly reducing the shudder that occurs
when a vehicle is started or accelerated.
In addition, our constant velocity joint design has
greatly reduced sliding resistance, and significantly
improved the idling vibration quality of A/T vehicles.
Retaining a mechanism similar to that of the PTJ,
our newly developed EPTJ features a more compact,
lighter CVJ design that has been realized through1. Introduction
Constant velocity joints (CVJ) used for automotive
drive shafts are categorized as fixed-type CVJs, whichenable a greater operating angle and are used on thetire side, and sliding-type CVJs, which are capable ofaxial sliding and are used on the engine side.
The shudder that occurs when a vehicle is started
or accelerated and the vibration that occurs during
idling of A/T cars are affected by vibration
characteristics that result from the induced thrust andslide resistance of sliding-type constant velocity joints.
In an attempt to solve these problems, in 2002, NTN
developes the revolutionary PTJ constant velocity joint
with vibration characteristic values that have been
reduced by more than 50% compared to conventional
low-vibration joints. Since then, the production of thisnovel constant velocity joint has been expanding.
With our PTJ products, the roller cassette in the
joint slides smoothly on the outer race roller groove
while maintaining specific attitude even when the joint
is operating at an angle, making it possible tosuppress the frictional resistance and reliably maintainthe vibration characteristic values at lower levels. As a

*Automotive Sales Headquarters

Fixed-type CVJ

Sliding-type CVJ

Front wheel drive axle

C.V.Joint Engineering Dept.

-80-

Compact and Shudderless Constant Velocity Joint (EPTJ)

optimization of the internal design and heat treatmentof joint components that fully utilize CAE techniques
including FEM and dynamic mechanism analysis, aswell as by adopting quality engineering (robust
design).

3. EPTJ functions
Making CVJs more compact and lightweight leads
to reduced vibration performance, mechanicalstrength and bearing life. We have overcome thesedrawbacks and our EPTJ has attained functionality
comparable to the PTJ.

2. EPTJ structure
The structure of our newly developed EPTJ is
illustrated in Fig. 1.
With an internal structure similar to that of the PTJ,
the EPTJ features lower frictional force throughimprovement in contact states and stabilization ofroller cassette attitude.
The internal structure of the EPTJ, being similar tothat of the PTJ, is structured as [roller cassette (inner
race/outer race/snap ring/needle rollers) + trunnion +outer race] to promote ease of assembly.
The outer race diameter of the EPTJ isapproximately 4% smaller than that of the PTJ (NTN
data: one size reduction) (Fig. 2). The weight of the
EPTJ is approximately 8% less than that of the PTJ
through size reduction of the outer race and innercomponents.

Conventional product:
PTJ

3.1 Reduction of induced thrust load(Countermeasure against accelerating
vehicle shudder)
Induced thrust load on a CVJ is an axial force
induced by mutual friction between joint components
when the shaft rotates at an angle with the outer race.
This thrust is associated with shudder in an
accelerating car, so reducing it can help suppress car
shudder.
Like the PTJ, the trunnion journal of the EPTJ is anelliptical cylinder (loading point: major axis side), andthe bore surface of the inner race in contact with thejournal is outwardly curved. Thanks to this contact
relationship, the moment of roller cassette tilt while the
CVJ is running is smaller and the attitude of the outer
race and roller cassette is stable. As a result, theinduced thrust on the EPTJ is always small regardlessthe operating angle as shown in Fig. 3. Incidentally, inthis diagram, the induced thrust values with the EPTJ
are nearly equivalent to those with the PTJ.

Newly developed product:
EPTJ

Induced thrust 3rd order component N▲4％Outer race
¡Outer race
¡Inner race
¡Sap ring×2¡Needle roller× n

Trunnion
Roller cassette×3

Fig. 1 Construction of PTJ and EPTJ

Outer race outside dia. mm

105

50

40

Competitor's low vibration CVJ

30

20

EPTJ
10

0
0

2

4

6

8

10

12

Operating angle deg.
PTJ
EPTJ

100
95

Fig. 3 Measurement of insulation resistance

90
85
80

EPTJ

75
70
65
70

60

PTJ
80

90

100

110

120

NTN joint size

Fig. 2 Outer-diameter comparison between PTJ and EPTJ

-81-

14

NTN TECHNICAL REVIEW No.73（2005）

3.2 Reduction in static plunging slide resistance(Countermeasure against D range idlingvibration with A/T cars)

4. FEM stress analysis and mechanicalanalysisSlide resistance on a CVJ, which occurs when theouter race is vibrating and allowed to slide in the axial
direction, is a cause of D range idling vibration in A/T
cars. Therefore, idling vibration can be suppressed bycontrolling the slide resistance.
Like the PTJ, the EPTJ is capable of oscillation on
the roller cassette bore and the trunnion journal. In
other words, with the EPTJ, the roller cassette rolls
only in the axial direction on the outer race, and theplunging force in the axial direction on the outer raceis borne only by the rolling motion of the needle
rollers, thereby keeping the slide resistance of the
EPTJ low regardless of the operating angle (Fig. 4).
Incidentally, in this diagram, the slide resistancevalues with the EPTJ are nearly equivalent to thosewith the PTJ.

To improve the overall strength, durability and NVH
characteristics in good balance, we determined the
optimal geometry of the major components (outer
race, trunnion and needle rollers) through full
utilization of FEM stress analysis. We sought lower
stress levels and compared evaluation results withactual components.
We have also fully utilized dynamic mechanism
analysis, as shown in Fig. 6, to understand theinternal forces and component motions that occur inoperation of the EPTJ in order to optimize its internal
dimensions.

（MPa）
500

Static plunging slide resistance N

100

300
80

Competitor's low vibration CVJ
100

60

40

EPTJ

-100

20

Fig. 5 FEM stress analysis for outer-race
0
4

6

8

10

12

Angle deg.

Fig. 4 Static plunging resistance vs operating angle
under torque and stimulated vibration condition

Roller cassette inclination

3.3. Strength
The mechanical strengths of the internal
components of the EPTJ, including static torsion
strength and torsion fatigue strength, are greater thanthose of outer race stems and shafts, meaning thatthe internal components are strong enough for use inreal cars.
Fig. 6 Dynamic mechanism analysis

3.4 Durability
The high and low load oscillation durability of the
EPTJ is equivalent to or better than that of the DOJ,
which boasts many years of good reputation. In
addition, the EPTJ durability is equivalent to or better
than that of a competitor's CVJ.
-82-

Compact and Shudderless Constant Velocity Joint (EPTJ)

6. Conclusion

The EPTJ is essentially a compact version of the
PTJ. Size reduction can cause problems including
decreased component rigidity and adverse effects on
bearing functions and characteristics due to theincreased contact pressure. In particular, the compact
design significantly affected the NVH characteristics,
and the induced thrust at a high torque load. To
address this problem, we applied quality engineering
to achieve a robust design.
In application of quality engineering, we first clarified
the functions of low vibration constant velocity joints,
and then listed the controllable factors, operating
conditions and variations in manufacturing that could
affect their functionality, before performing anexperiment using the L18 orthogonal matrix. We
selected the factors that significantly affect thefunctions of constant velocity joints and adopted our
findings in designing the EPTJ.
As shown in Fig. 7, the induced thrust on the EPTJ
was decreased by approximately 50% throughapplication of quality engineering.

EPTJ features:(1) The magnitude of induced thrust is not affected by
the operating angle of the joint, and the EPTJ can
attain low vibration comparable to the PTJ.
(2) The magnitude of static plunging slide resistance is
not affected by the operating angle of the joint, andthe EPTJ can attain low vibration comparable to the
PTJ.
(3) The outside diameter of the outer race isapproximately 4% smaller than the PTJ (NTN data:
one size reduction).
(4) The weight is approximately 8% lighter than the PTJ.
(5) The maximum operating angle is the same as that of
the PTJ (26˚).
The EPTJ is a low vibration CVJ that is muchsmaller and lighter compared with conventional CVJs
and that is minimally affected by the operating angle.
EPTJ units will fully demonstrate their design
performance in conditions when better space
efficiency is needed and a larger operating angle isrequired due to the layout limitations of a particularcar, and when greater torque is applied to the CVJ
when the car is idling, starting or accelerating.

Induced thrust 3rd order component N

Induced thrust 3rd order component N

5. Application of quality engineering

150

Before application ofquality engineering
Load torque A
Load torque B

100

Load torque C
Load torque D

50

0

0

5

10

150

After application ofquality engineering
Load torque A
Load torque B

100

Load torque C
Load torque D

50

0
0

Operating angle deg.

5

Fig. 7 Quality engineering for induced cyclic load reduction
Photos of authors

Tatsuro SUGIYAMA

Yuuichi ASANO

Automotive Sales Headquarters
C.V.Joint Engineering Dept.

Automotive Sales Headquarters
C.V.Joint Engineering Dept.

10Operating angle deg.

-83-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Constant Velocity Steering Joint (CSJ)

Kenta YAMAZAKI**

In an automotive steering system, Cross or Cardan joints typically were used. However, because of
the nature of non-constant velocity joints, we are required to use two joints with special geometricalconstruction to obtain constant velocity performance. To improve this situation, a new constantvelocity steering joint (CSJ) has been developed.
This paper reports the summary of the concept, design and characteristics of this new joint
developed by NTN. Adding the original idea to the reliable know-how of constant velocity universal
joint, a smooth and non-backlash feeling has been achieved.

shafts and in unique constructions, NTN has become
the first in the world to develop a CSJ constantvelocity ball-type steering joint that is capable ofsmooth, constant velocity actuation over the entire
operating range. This report describes the features of
our CSJ and the results of basic functionality tests.

1. Introduction
Cross joints, also known as Cardan joints, (see Fig.
8) have been used for automotive steering systems.
Since a cross joint is not a constant velocity joint, if
only one cross joint is used to provide a larger
operating angle for an automotive steering system,
then variation in rotational angle and resultant
variation in torque occurs, leading to uncomfortablesteering. To avoid this problem a feature similar to thatobtained with constant velocity joints can be achieved
if two joints are combined such that their operating
angle is the same. They are positioned in rotational
phases where the deviation in rotational velocity of
one joint is counterbalanced by the other joint.
With recent automotive models, the available spacein engine compartments is becoming smaller andsmaller due to the conflicting needs for both compact,
lightweight car bodies and greater passenger space.
As a result, there has been increased difficulty inrealizing two-joint system steering layouts that provide
constant velocity features.
Using our expertise with constant velocity joints thathave performed outstandingly with automotive drive*Automotive Sales Headquarters

2. Conventional technology
Steering joints are expected to operate smoothly
and free from backlash in the rotational direction,
while greater operating angles and compactness aredesired. Currently double Cardan joints (D-CJ) (seeFig. 9) are mass-produced as constant velocity
steering joints that consist of two cross joints with acentering mechanism so that one D-CJ unit canfunction as a constant velocity joint. However, thedouble Cardan joints have problems including greater
size, heavier weight and low torsional rigidity. D-CJs
can be categorized into two types: pseudo constantvelocity and full constant velocity. Both types havedrawbacks, including unreliable constant velocity with
angles outside the preset operation angle range and anarrower oscillating angle (see Table 1).

C.V.Joint Engineering Dept.

-84-

Constant Velocity Steering Joint (CSJ)

With this structural arrangement, even when minor
wear occurs on contact areas between thecomponents due to extended steering system use,
associated clearances are automatically reduced
through the action of the spring, maintaining thesteering system in a backlash-free state.
NTN has also attempted to optimize the shapes of
the components and the clearance between them, andhave optimized the force of the spring accordingly. In
this way, we have achieved both reduction inrotational backlash throughout the operating angle
range and smoother steering system rotationalmotion.

3. CSJ structure
The basic structure of the CSJ resembles that of the
BJ (fixed constant velocity joint of track offset type)that is often used for drive shafts. However, with an
ordinary BJ, clearance between the ball groove(formed between the outer race and inner race) andthe balls is small, and as a result, backlash occurs in
the rotational direction. Because of this, the BJ is notsuitable for automotive steering systems that need tobe free from backlash. Another problem is that if atightening allowance is provided for clearance
between the ball groove and the balls in order toreduce backlash, assembly of the BJ into the steeringsystem may not be possible.
To solve these conflicting problems, we have added
a plunger and a spherical plate to the joint (see Fig.
1). More specifically, a plunger comprised of a spring
and a ball is provided at the shaft end, and the force of
this spring is applied through the ball to the sphericalplate fitted into the cage. As a result, as shown in Fig.
2, the inner race alone shifts to the right, causing theball to be lifted until it touches the outer race, thereby
the clearance on the ball groove is reduced.

4. Performance characteristics
4.1 Rotational backlash
The results of comparison of our CSJ with a D-CJ
are illustrated in Fig. 3.
Since it is free from excessive backlash, our CSJ
boasts higher torsional rigidity, compared to currently
used double Cardan joints (D-CJ), for automotive
steering systems.

CSJ

Inner race
Shaft

Band
Outer race
Spherical plate
PlungerBand+
Cage Ball

CSJ
D-CJ

Boot

Fig. 1 Structure

Torsional torqueSpherical plate0

Displacement inaxial direction

Plunger--0

Torsional angle

Fig. 3 Rotational back lashFig. 2 Structure to reduce rotational play

-85-

+

NTN TECHNICAL REVIEW No.73（2005）

4.2 Rotational torqueOperating time

Test result

Targeted life Targeted life×2

The results of the comparison of the CSJ and the DCJ are summarized in Fig. 4.
Though somewhat greater than that of the D-CJ, therotational torque of the CSJ falls in a permissible
range and has less variation. Therefore, these CSJ
characteristics are suitable for steering joints.

Condition A

No problem

Condition B

No problem

Fig. 5 Endurance test results+
CSJ
Torsional torque

Rotational torque

D-CJ

0
0

60

120

180

240

300

No failure
Development target

±0
103

360

Rotational angle deg.

104

105

106

107

Number of twists

Fig. 4 Rotating torqueFig. 6 Torsional fatigue strength

5. Endurance and strength

Development target

Failure torqueThe results of an endurance test are summarized inFig. 5, and the results of strength tests are plotted in
Figs. 6 and 7.
In terms of both endurance and strength, the CSJ
satisfies targeted levels, and we believe that the CSJ
can be used in applications on actual automobiles.

6. Comparison with other constantvelocity steering joints
Compared with competitors' double Cardan joints,
the CSJ is lighter and more compact, and has agreater permissible operating angle. Therefore, theCSJ can greatly improve the freedom of layout for
automotive steering systems.

0

φ64 mm

φ61.5 mm

φ70 mmWeight

441ｇ

970ｇ

830ｇ

Volumeoccupied

160 cm3

327 cm3

270 cm3

0〜48 deg
Operating (Fully constantangle
velocity overwhole range)

0〜47 deg

40〜48 deg

(Pseudo constantvelocity type)※(Fully constantvelocity type)

20

30

40

Fig. 7 Static torsion strength

Double Cardan joint Double Cardan joint
from competitor A from competitor B

Outside dia.

10

Joint angle

Table 1 Comparison of constant velocity steering joint
CSJ

0※Constant velocity feature is lost when the angle is outside thepermissible range.

-86-

50

Constant Velocity Steering Joint (CSJ)

7. Conclusion
Developed as a dedicated constant velocity steering
joint for automobiles the NTN CSJ is characterized bya unique preload applying mechanism incorporated
into the joint and internal clearance optimized torealize not only inhibition of rotary-direction backlash
but also the capability for smooth rotational operation
over the entire operating angle up to the maximum
operating angle of 48 degrees. These features, which
are lacking in conventional BJs, make the CSJ
capable of providing an ideal constant velocity for
automotive steering systems, while maintaining
sufficient endurance and mechanical strength.
With the CSJ, one joint unit provides smooth,constant velocity transmission of rotation, adjustmentof joint angles and phase matching in the rotational
direction. In contrast, two Cardan joints are needed forthis application. Furthermore, compared with double
Cardan joint arrangements, the CSJ features lighter
weight and a more compact design, allowing
significantly greater freedom for steering systemdesign layout. We believe that our CSJ is the constantvelocity joint type that can satisfy mounting needs foradvanced functionality and more compact size inautomotive steering systems.

1
3
2

1 Bearing
2 Cross spider
3 Bearing seal
4 Yoke4

Fig. 8 Crose joint (Cardan joint)

SocketSocket yoke
Centering ball
Needle rollers

Outer seal
Inner seal

Inner seal
Bearingcup
Coupling yoke
Bushing
Socket yoke

Needlerollers

Reference
1) The Society of Automotive Engineers of Japan,
Automotive Engineering Handbook, Design Volume(1991)

SpiderFig. 9 Double cardan joint

Photos of author

Kenta YAMAZAKI
Automotive Sales Headquarters
C.V.Joint Engineering Dept.

-87-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Summary of E Series, Constant Velocity Joint
for Propeller Shaft

Tomoshige KOBAYASHI**
Teruaki FUJIO**

Automotive manufacturers are trying to develop automobiles with better fuel efficiency and comfort.
NTN developed E series CVJs for propeller shaft based on the design of mass-produced E series
high efficiency constant velocity joints for drive shafts.
This paper summarizes the E series CVJs for propeller shaft products.

1. Introduction

2. Overview of E Series CVJ products
for propeller shafts

To satisfy automotive performance requirements,
including higher efficiency and lower vibration, NTN
has been mass-producing E series constant velocity
high-efficiency joint products for drive shafts. The
need has been mounting to satisfy these requirements
for constant velocity joints for propeller shafts incomplicated 4WD car drive systems. To address thistrend, we have developed E series constant velocity
joints (CVJ) for propeller shafts. The features ofvarious E series CVJ subtypes are described below.

We have developed high-speed variants of the Eseries -the HEBJ, HEDJ and HETJ- whose basic
structure derives from that of other E series drive shaft
products. Compared with conventional massproduced CVJs for propeller shafts (BJ, DOJ and TJ),
the HEBJ, HEDJ and HETJ series products boast highefficiency, light weight and compact size while
maintaining the load carrying capacity and durability of
the conventional products. The features of the Eseries CVJs for propeller shafts are summarized inTable 1.
Incidentally, in consideration of the environment,
NTN uses a lead-free grease prepared especially forpropeller shaft CVJs and employs a hexavalent
Table 1 Characteristics of E series
New
Maximum Outside
Torque Heat
Subject ofMassseries allowable angle dia.
loss generation comparison

Photo 1 Example of Propeller shaft. (2 joint design)*Automotive Sales Headquarters

HEBJ

＋

＋＋＋ ＋＋＋ ＋＋

＋＋

BJ

HEDJ

＋

＋＋

＋＋

＋＋

＋＋

DOJ

HETJ

＋

＋＋ ＋＋＋＋＋＋

TJ+++ …Excellent

C.V.Joint Engineering Dept.

-88-+++… Good+… EquivalentSummary of E Series, Constant Velocity Joint for Propeller Shaft

Table 2 Comparison of outer-diameter between
HEBJ and BJ

chromium-free material for CVJ components that needsurface treatment.

Unit: mm

NTN joint size

3. E Series Propeller Shaft CVJ productfeatures and future developments

BJ

HEBJ

87

79.9

72.6

100

91.0

83.7

Flange type design of Constant Velocity Joint for Propeller Shaft wascompared with the outer-diameter.

3.1 HEBJ
3.1.1 Features
When a CJ (cross joint) cannot satisfy requirements
for a greater operating angle, a BJ type CVJ for
propeller shafts is used together with a sliding type
CVJ.
Like the EBJ for drive shafts, the HEBJ for propeller
shafts features temperature increase and transmission
efficiency that is much improved over conventional BJ
products.
The construction of HEBJ is illustrated in Fig. 2. A
size comparison of the HEBJ and a conventional BJ
for propeller shafts is provided in Fig. 2, and theiroutside diameters are shown in Table 2.

The flange type and disk type designs are available
for the HEBJ in order to cope with user requirements.
(See Figs. 3 and 4.)Outer race
Cage

Ball

Fig. 3 HEBJ flange type design

Inner race

Fig. 1 Composition of HEBJ

Fig. 4 HEBJ disc type design

3.1.2 Future development

BJ

We will make HEBJ products the standard bearings
for fixed constant velocity joints for propeller shafts,
and will increasingly replace conventional BJ and CJ(cross joint) products with HEBJ products.

HEBJ

Fig. 2 Comparison of structure between BJ and HEBJ

-89-

NTN TECHNICAL REVIEW No.73（2005）

3.2 HEDJ
3.2.1 Features

3.3 HETJ
3.3.1 Features

The DOJ type CVJs for propeller shafts are capableof a greater axial slide, and can absorb minute axial
vibration.
Like the EDJ products for drive shafts, thetemperature increase on the HEDJ is limited and an
induced thrust on the HEDJ is lower compared withthe conventional DOJ. The size of the HEDJ isapproximately 4% smaller than that of the DOJ. To
cope with varying CVJ use specifications for propeller
shafts, the HEDJ is available in sizes #71, #82, #92
and #95. The standard HEDJ forms available are thepressure-welding type and flange type, both of which
are in demand.
The joint construction (pressure-welding type) is
illustrated in Fig. 5, and a comparison of outside
diameters with the DOJ is given in Table 3.

The permissible slide amount is greater with TJ type
constant velocity joints for propeller shafts. TJ
products feature smaller slide resistance over thewhole sliding range as well as lower heat generation.
As for the ETJ for drive shafts, the internal design of
the HETJ has been improved based on data obtained
from FEM analysis. As a result, the HETJ products forpropeller shafts feature a smaller outside diameter,while their strength, durability and NVH characteristics
are equivalent to those of the conventional TJ.
The standard form is a pressure-welding type.
The construction of the joint is illustrated in Fig. 6, a
comparison of proportions of the TJ and HETJ for
propeller shafts is given in Fig. 7, and a comparison ofoutside diameters of the TJ and HETJ is given in
Table 4.

Cage

Needle roller
Outer race

Outer race
Ball

Inner race

Trunnion
Spherical rollers

Fig. 5 Composition of HEDJ

Fig. 6 Composition of HETJ

Table 3 Comparison of outer-diameter between
HEDJ and DOJ
Unit: mm

NTN joint size

DOJ

HEDJ

71

―

63

75

73

―

82

―

73

87

79

―

92

―

79

95

―

83

TJ

HETJ

Fig. 7 Comparison of structure between TJ and HETJ

-90-

Summary of E Series, Constant Velocity Joint for Propeller Shaft

Table 4 Comparison of outer-diameter between
HETJ with TJ

4. Conclusion

Unit: mm

NTN joint size

TJ

HETJ

75

68

66

79

73

68

By applying the experience and expertise
accumulated through the development of the E series
CVJs for drive shafts to CVJs for propeller shafts, NTN
has successfully developed the lightweight, compact E
series CVJs for propeller shafts.

3.4 Future development of sliding types

NTN will further improve the CVJs for propeller it ispossible shafts, for which demand will increase, sothat to supply constant velocity joints that best satisfy
the customer needs.

An optimal choice can be selected from the sliding
types (HEDJ and HETJ) and other sliding joint types(HLJ and LJ), according to the required slide amount,
outside diameter size and CVJ characteristics (seeTable 5).

Table 5 Comparison of characteristics between propeller shaft CVJ
Fixed typeSliding type

HEBJ

HEDJ

HETJ

HLJ

LJ

Slide amount

―

◎

◎

○

△

Permissible angle

◎

◎

△

○

◎

High speed operation

○

○

△

◎

◎

Heat generation

○

○

◎

△

△

NVH Slide resistancecharacteristics Induced force

―

○

◎

△

△

Backlash

8th order

8th order

3th order

6 and 3th order

6 and 3th order

Radial direction

―

○

△

◎

◎

Rotation direction

○

○

△

◎
◎：Excellent

Photos of authors

Tomoshige KOBAYASHI

Teruaki FUJIO

Automotive Sales Headquarters
C.V.Joint Engineering Dept.

Automotive Sales Headquarters
C.V.Joint Engineering Dept.

-91-◎○：Good△：Average

NTN TECHNICAL REVIEW No.73（2005）[ Technical Paper ]NVH Analysis Using Full Vehicle Multi Body Dynamic Model*-Influence of Constant Velocity Universal Joints on Shudder Vibration–

Yoshihiko HAYAMA****
Takashi NOZAKI****
Masafumi NAKAKOUJI****
Satoshi FUJIKAWA****
Komaki FUKUSHIMA****

Vehicle NVH characteristics have been analyzed using a full vehicle simulation model. This model
consists of engine, drive train and chassis components that are respectively accurate multi body
dynamic models. This paper focuses on a shudder vibration analysis, specifically on the influence ofconstant velocity universal joints (CVJs) on shudder vibration characteristics.

analysis have been reported. 2) Analysis techniques
are used not only for vehicle development but also for
automotive component development. 3) 4)

1. Introduction
In developing automobiles, the mode of NVH
problems varies significantly depending on engine andchassis types, with many problems occurring only incompleted vehicles. In addition, the NVH problems tobe solved during development of a particular car
require many hours of work and pose a significant
challenge in expediting vehicle design completion.
Furthermore, since the NVH problems of a given
vehicle are often the result of complex mechanisms,
determination of the root causes of each NVH problem
through experiments with actual vehicles is very
difficult.
Recently, various structural and mechanism
analysis techniques have been applied to vehicle
development, 1) contributing to the simplification of
prototyping and reduction of labor time required forvehicle development. As for NVH problems, examples
of analysis techniques applied to idling vibration

We have developed a full vehicle dynamic modelusing mechanism analysis to analyze vehicle
vibration. The NVH problem we analyzed this time
was the shudder vibration that occurs when a vehicle
starts and runs straight ahead. This full vehicle
dynamic model is a detailed mechanism analysis
model that incorporates virtually all the elements
usually included in automobiles, including a power
train ranging from the engine to tires as well as asteering and suspension system. Using this fullvehicle model, we have reviewed the effects of a drive
shaft constant velocity joint on shudder vibration. The
constant velocity joint model also reflects nonlinear
elements in its mechanism simulation. We haveverified that this model reproduces the forces that act
on the interior of actual constant velocity joints. 3) 4)****Published in the Spring Scientific Lecture Session, the Society of Automotive Engineers of Japan, May 19, 2004****New Product Development R&D Center New Product Development Dept.
****Automotive Sales Headquarters Automotive Engineering Dept.
****Mazda Motor Coproration Powertrain Advance Development Dept.

-92-

NVH Analysis Using Full Vehicle Multi Body Dynamic ModelAs shown in Fig. 2, the major component of the
shudder vibration is clearly the drive shaft rotation 3rdorder. The drive shaft of the test vehicle is equippedwith a tripod type constant velocity joint (hereafter, TJ)
on its inboard side and a fixed constant velocity joint(hereafter, EBJ) with eight balls on its outboard side.
Because of this, it appears that the inboard side
constant velocity joint is responsible for shudder
vibration on starting vehicles.
A non-mass-production sliding constant velocity
joint was used in our experiment in order to analyze
the shudder vibration phenomenon accurately.

2. Shudder vibration in starting
vehiclesShudder vibration occurring in a starting vehicle isgenerally defined as the phenomenon in which a car
body is excited by resonance between the engine
vibration and the induced thrust on a sliding constantvelocity joint.
Results of shudder vibration measurements
obtained on an actual vehicle are summarized in Fig.
1, where Fig. 1 (a) illustrates vehicles speeds, Fig. 1(b), torques acting on the drive shaft, and Fig. 1 (c),lateral acceleration below the driver's seat floor.
The area encircled with a dotted line represents astate where lateral vibration occurred and the vehicle
accelerated while vibrating left and right. The
conditions that induced shudder on the test vehicle
were a time of 2.5 to 3.5 seconds after starting,vehicle speed of 28 to 40 km/h, and drive shaft torqueof approximately 700 Nm, and the associated drive
shaft speed was in a range of 230 to 330 rpm.
Fig. 2 illustrates the results of tracking analysis withthe drive shaft torque values provided in Fig. 1 (c).

Speed km/h

60
50
40
30
20
10
0

Acceleration
m/sec2

1

0

250

300

350

Fig. 2 Tracking result of fig.1 (c)3. Full vehicle model
3.1 Power train and chassis model

0

1

2
3
Time sec

4

5

4

5

4

5

To theoretically forecast the shudder vibration
phenomenon, analyses of engine vibration
characteristics and induced thrust on a constantvelocity joint, as well as transmission of the engine
vibration and the induced thrust to the car body via themount and suspension, are also necessary. To thisend, we developed a detailed full vehicle mechanism
simulation model and performed analysis by applying
a drive torque from the engine output in order tosimulate an induced thrust occurring from rotation of
the constant velocity joint.
Fig. 3 schematically illustrates the power train andchassis model used for our analysis work. The engine
model simulated the piston and crank motion resulting
from the pressure of the fired fuel. The model for thetorque-transmitting transmission includes thesimulated torque amplification by a torque converter,
speed reduction by planetary gearing and motion of adifferential. The model for suspension includes aspring mass capable of simulating the suspension
resonance characteristics, and a suspension masscapable of simulating the rigidity of the stoppers on astarting vehicle. Furthermore, the tire model accurately
simulates tire rigidity, thus allowing the simulation of acomplete vehicle running on a road, including friction
with the road surface.

1500
Torque Nm

200

Drive shaft speed rpm(a) Vehicle speed

1000
500
0
0

1

2
3
Time sec(b) Drive shaft torque1
Acceleration
m/sec2

Over all
Drive shaft 3rd order

0

-1
0

1

2
3
Time sec(c) Acceleration of floor(Lateral direction; 50Hz LP Filter)
Fig. 1 Experimental Results of Shudder Vibration-93-

NTN TECHNICAL REVIEW No.73（2005）(a) Inboard CVJ (TJ)

Fig. 3 Multi body Powertrain & Chassis model

Our vehicle model was developed withconsideration for not only shudder vibration but also
possible application for evaluation of other vehicle
NVH performance. Thus, the vehicle model includes
detailed simulation of sections not associated withprediction of shudder vibration.

3.2 Drive shaft model(b) Outboard CVJ (EBJ)As described in Sec. 2 above, we learned that thedrive shaft 3rd order component affects shudder
vibration. The test vehicle has the TJ on its inboard
side. In order to simulate the drive shaft 3rd order, theconstant velocity joint model must be a detailed
mechanism simulation model. Fig. 4 shows theconstant velocity joint models (TJ and EBJ) that wereused in our full vehicle dynamic model. Thus, thisconfiguration provided mechanism simulation-capableinboard and outboard constant velocity joint models.
Each of these mechanism models simulated theclearance, contact force and friction between therelated components in the CVJs. 3) 4)

Fig. 4 Multi body CVJ model

Acceleration
m/sec2

1

0

-1
200

250

300

350

300

350

Drive shaft speed rpm

Acceleration
m/sec2

1

4. Verification of analysis accuracy

Drive shaft 3rd order
Experiment
Simulation

0200

Fig. 5 shows a comparison between the analysis
results for lateral acceleration below the floor at thedriver's seat and the actual experiment results.
Though the amplitude obtained from the simulation issmaller than that obtained from the experiment, theresults of analysis match those of the experiment well.

250

Drive shaft speed rpm

Fig. 3 Acceleration of floor (Lateral direction)

range of approximately 11.5 to 16.5 Hz. To investigate
the vehicle vibration mode in this frequency range, we
performed model analysis with the full vehicle model.
Fig. 6 illustrates the lateral direction engine mode.
The characteristic frequency of the lateral engine
motion was approximately 15 Hz. The drive shaft
speed was approximately 310 rpm when theacceleration was greatest as shown in Fig. 5, and thefrequency of the drive shaft rotation 3rd ordercomponent was approximately 15.5 Hz.

5. Analysis of shudder vibration
phenomenon
5.1 Relation between engine and drive shaft
We found that shudder vibration on a starting
vehicle is related to the drive shaft rotation 3rd ordercomponent. The speed of a drive shaft that triggers
shudder vibration is in the range of 230 to 330 rpm.
The frequency of the 3rd order component falls in a-94-

NVH Analysis Using Full Vehicle Multi Body Dynamic Model

Because of this, shudder vibration seemed to occur
from the resonance between the lateral acceleration of
the engine and the drive shaft rotation 3rd ordercomponent.
For further clarification, we altered the rigidity of theengine mount in the mechanism simulation model,
and again performed modal analysis and lateral
shudder analysis. Fig. 7 shows an engine vibration
mode with reduced engine mount rigidity. The
alteration to rigidity can be understood to change thecharacteristic vibration frequency and mode. Fig. 8offers a comparison of the magnitudes of shudder
vibration between these two modes. In summary, foreach particular engine mode, there is a uniquevibration frequency and a unique magnitude of
shudder vibration.

Induced thrust, which is one of the drive shaft
rotation 3rd order components, typically occurs with
the TJ. When the TJ is run, an induced thrust occurs
along the rotational axis of the outer race and thatmainly consists of the rotational 3rd order. This
induced thrust is a force that results from frictional
forces occurring because of the mutual friction of therotating internal components on the TJ.5) Therefore,
we executed shudder vibration analysis with two
sliding constant velocity joints with unique friction
characteristics. Fig. 9 summarizes the results ofanalysis for lateral acceleration below the floor at thedriver's seat. As can be seen from this diagram, it
seems that use of a low friction CVJ can alleviate
shudder vibration.

Fig. 6 Lateral engine mode shape (14.93Hz) of originalmodel

Fig. 7 Lateral engine mode shape (12.33Hz) of modified
model

5.2 Influence of CVJ type1

Acceleration
m/sec2

Acceleration
m/sec2

1

Drive shaft 3rd order
Original
Modified0

Drive shaft 3rd order
CVJ A
CVJ B
0

200

250

300

350

200

Drive shaft speed rpm

250

300

Drive shaft speed rpm

Fig. 8 Effect of engine mode on shudder vibration

Fig. 9 Effect of CVJ on shudder vibration(Induced Thrust CVJ A>CVJ B)

-95-

350

NTN TECHNICAL REVIEW No.73（2005）

is apparent from Fig. 12, this trend was validated in
the experiment with an actual vehicle. Consequently,
we learned that this variation in shudder vibration leveloccurred when the differential gearing was activatedas the vehicle turned and, as a result, phase gap
between the right hand and left hand TJs varied.

5.3 Effect of phase gap with right-hand andleft-hand CVJs
The shudder vibration test with an actual car was aforward traveling-only starting and acceleration test.
When the vehicle turned, the resultant measurement
values greatly varied. Fig. 10 illustrates the running
pattern used in the shudder test and the resultant
maximum measurements of shudder vibration. The
measurement operation was repeated 30 times foreach constant velocity joint type. Each CVJ typeexhibited measurement variation and had uniquevibration characteristics. Notwithstanding, the lowest
shudder measurements for both CVJ types were
nearly same.
There are right-hand drive and left-hand drive
shafts. The right hand and left hand TJs areincorporated via the differential box. The resultant
force of the induced thrusts from the two TJs varies
depending on the phase gap between both TJs. Fig.
11 (a) shows the definition of phase gap, and Fig. 11(b) plots induced thrusts resulting from the phase gap
angles between the two TJs. As can be seen in Fig.
11(b), when the relative phase of one TJ coincides
with that of the other TJ, the resultant force of theinduced thrusts theoretically becomes zero, andreaches the maximum level when the phase gap
angle is 60 degrees.
With our full vehicle model, the effect of the phasegap between the right-hand TJ and the left-hand TJ
was analyzed. As shown in Fig. 12, it was foundthrough simulation that the magnitude of shudder
vibration was largest when the phase gap was in arange of 45 to 60 degrees and was smallest when thephase gap was in a range of 100 to 120 degrees. As

Left Hand TJ

Right Hand TJ
Trunnion
Tr
Axis (RH)
Ax unniis ( onPhase
gap LH)

Dash Line:
Right Hand TJ

Solid Line:
Left Hand TJ

Front Sideof Vehicle

Rear Sideof Vehicle

3rd order compornent
of induced thrust N(a) Definition of phase gap

0

30

60

90

120

Phase gap angle deg.
(b) Induced thust
Stop

Takingoff

Fig. 11 Theoretical effect of CVJ phase gap
on induced thrust
Braking

Acceleration
Turn

Turn
Braking

Acceleration
1Stop(a) Running pattern

0

1

Acceleration
m/sec2

Experiment
Simulation

Acceleration
m/sec2

Takingoff

Drive shaft 3rd order
CVJ A
CVJ B

30

60

90

120

Phase gap angle deg.

Fig. 12 Effect of CVJ phase gap on shudder vibration

0

CVJ A

0

CVJ B

(b) Experimental data (30 times par each CVJ)

Fig. 10 Mesurement summary of shudder test-96-NVH Analysis Using Full Vehicle Multi Body Dynamic Model

6. Conclusion
We developed a full vehicle running-capablemechanism simulation model and performed asimulation for vehicular NVH (analysis of shudder
vibration).
We identified and verified the relationship between
the shudder vibration mechanism on a starting vehicle
and engine vibration modes.
We qualitatively and quantitatively clarified theinterrelation between induced thrusts and shudder
vibration levels on constant velocity joints.

References
1) Kurisu, Fujikawa, Miyauchi, Koizumi, Hirobe,
Fukushima: Application of ADAMS mechanism
analysis software to power trains, Mazda Technical
Review, No. 22, p44-49 (2004).
2) Matsumoto, Suzuki, Kitani, Kanuma: Idling shudder
analysis using a full vehicle model, the Society ofAutomotive Engineers of Japan, 2003 Spring
Scientific Lecture Session preprints, No. 39-03, p1-4
(2003).
3) Hayama: Dynamic internal force analysis for constantvelocity joints considering nonlinear elements, theJapan Society for Computational Engineering andScience, lecture proceedings, Vol. 8, No. 1, p393396 (2003).
4) Hayama: Internal force analysis for constant velocity
joints, the Society of Automotive Engineers of Japan
Journal, Vol. 34, No. 4, p157-162 (2003).
5) Nozaki, Kobara: Experimental determination of forces
acting on tripod constant velocity joint housings, theJapan Society of Mechanical Engineers, 2003 annual
meeting proceedings, IV, p155-156 (2003).

Photos of authors

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

Masafumi NAKAKOUJI

Satoshi FUJIKAWA

Komaki FUKUSHIMA

Automotive Sales Headquarters Automotive Sales Headquarters
Automotive Engineering Dept.
Automotive Engineering Dept.

Takashi NOZAKI

Mazda Motor Coproration
Powertrain Advance
Development Dept.

Mazda Motor Coproration
Powertrain Advance
Development Dept.

-97-

NTN TECHNICAL REVIEW No. 73（2005）[ Technical Paper ]Development of a High Precision Angle Sensor

Toru TAKAHASHI**
Yoshitaka NAGANO**
Shoji KAWAHITO**

Small and high-resolution encoders capable of absolute angle detection are required for many applications
such as automotive, industrial, consumer products, and robots. For automotive applications, these sensors areused not only for the motion control of mechanical parts and motors, but also for sensing the human operating
motion such as steering sensors or pedal sensors. However, commonly used rotary encoders using magnetoresistive devices or Hall devices combined with a multi-pole magnet have difficulty in meeting theserequirements because they have a large cylindrical structure and need a fine-pitched magnetic scale.
Furthermore, they can not detect absolute angles. This paper presents a single-chip CMOS magnetic rotary
encoder system that operates on a new working principle based on magnetic pattern analysis with a statistical
calculation method. The developed chip can detect the absolute angle at a resolution of 10bits per rotation,
which leads to a promising solution for compact high-resolution rotary encoders.

detect magnetic field variations. Since they are lesssusceptible to the effects of dust, magnetic rotary
encoders are used more often in automotive
applications.
On many commonly used magnetic rotary
encoders, a magnetic sensor, such as a Hall sensor oran MR device, is used to detect the magnetic field
variation that results from rotation of a multi-polemagnet. Most of the NTN bearings equipped withrotary encoders use this type of a magnetic sensor.
However, this arrangement poses an obstacle tocreating more compact rotary encoders because thepole pitch of the multi-pole magnet would be too small
for the magnetic sensor to detect magnetic field
variations. Therefore, a new sensor that operates on adifferent principle has been needed to provide acompact detection system that is capable of highprecision angle detection.
In cooperation with Professor Shoji Kawahito of the
Shizuoka University Research Institute of Electronics,

1. Introduction
Rotary encoders for rotational angle detection havebeen used in a variety of industrial applications,
including detection of actuation and travel amounts
and motor speed control. In automotive applications,rotary encoders are utilized for steering system angle
detection, accelerator pedals and throttle valves toimprove vehicle operability and safety. For every
automotive application, the need has been mounting
for more compact rotational angle detection sensors
that are capable of higher detection accuracy and areless expensive.
The detection schemes of rotary encoders can be
categorized as contact and non-contact types.
Components on a contact type rotary encoder candeteriorate from wear. Non-contact rotary encoders
can be further subcategorized into optical rotary
encoder types that utilize transmission or reflection oflight beams and magnetic rotary encoder types that***New Product Development R&D Center Mechatronics Research Dept.
***Shizuoka University Research Institute of Electronics Professor-98-Development of a High Precision Angle Sensor

NTN has developed a unique compact magnetic arrayrotary encoder that measures 5 mm square and iscapable of detecting an absolute angle of ±0.36˚. At
the International Solid-State Circuits Conference 2005
(ISSCC 2005) held in February 2005, this sensor won
the Beatrice Winner Award for Editorial Excellence.1)
This paper provides an overview of our sensor andreports the results of angle detection with a compact
bearing that incorporates this sensor.

signals, an analog-to-digital converter (ADC) circuitthat converts analog signals into digital signals and adigital processor circuit that calculates rotational
angles. This integrated sensor chip has all thefunctions needed to detect the absolute rotational
angle of a magnet and output detection results asnumerical data.

2.2 MAGFET magnetic sensor element
Our sensor chip uses a split-drain MAGFET(Magnetic Field Effect Transistor) for each magnetic
sensor element. MAGFETs can be manufactured by a
CMOS (Complementary Metal-Oxide Semiconductor)
process similar to that used for ordinary logic circuits.
They can also be incorporated into chips that also
include amplifier and calculator circuits.
The MAGFET structure and corresponding circuit
are illustrated in Fig. 3.
Like common transistors, each MAGFET has four
terminals, including a source (S), a gate (G) and two
separate drain (D) terminals. This sensor is sensitive
to magnetic fields that are perpendicular to the sensor
elements, and functions because of the Lorentz force
used for the current within the transistor. When amagnetic field is not present, the current levels of thetwo drain terminals are the same and the relation I1=I2
holds. When a magnetic field Bz is applied, electrons
will flow in the direction of the red arrow in Fig. 3, andthere will be a bias between the currents of both drain
terminals, resulting in a difference between the draincurrents (I1-I2) that occur in accordance with themagnetic field applied.
When assuming that the magnetic field intensity is
Bz [T], the sensor sensitivity is SR [1/T] and the sensor
bias current holds a relation of lbias = I1+ I2, then thesensor output signal can be defined as follows:

2. Magnetic array sensor operating
principles
2.1 Structure of sensor
The rotary encoder we have developed consists ofa permanent magnet (using sintered NdFeB) at the tipof a rotary shaft and a 5 mm square sensor chip
opposite the permanent magnet with a gap of
approximately 0.5 to 1 mm between them (Fig. 1).
The internal configuration of the sensor circuit
formed on the sensor chip is schematically illustrated
in Fig. 2. The integrated sensor chip is comprised of amagnetic sensor array (MAGFET Array) consisting ofmultiple miniature sensor elements arranged along thefour edges of a tetragon. These include a sample-andhold amplifier (SHA) circuit that amplifies sensor

Isignal = I1−I2= SR BZ Ibias……………………………（1）Fig. 1 Rotary encoder setup

Fig. 2 Block diagram of the sensor chipFig. 3 Structure of MAGFET (unit μm)-99-

NTN TECHNICAL REVIEW No. 73（2005）

2.3 Angle detection scheme

Sensor signal

The arrangement of the sensor arrays is shown in
Fig. 4. This diagram illustrates a state where themagnet is at a standstill with a rotational angle θ. The
blue line in this diagram shows the border between
the N-pole and the S-pole. The magnetic field
perpendicular to the page from this line is zero, so thisline is called the "Zero-Line."
When the sensor signals from the multiple
MAGFETs arranged on the sensor chip aresequentially read in a clockwise direction, beginningwith the No. 1 position in the upper left corner of Fig.
4, and the read signals are plotted, the wave patternshown in Fig. 5 is obtained. The relation between themagnetic field distribution and the sensor signals isshown in Fig. 6. By detecting two points, Z1 and Z2,
where the obtained sensor signal level is zero, theangle of the Zero-Line can be calculated based on
their coordinates.

Boundary
between poles

N-poleS-poleTwo points where no magnetic field exists are detected.
→ Boundary angle is calculated.

Fig. 6 Detection of the Zero-Line

To improve the detection accuracy of two zerocrossing points in the actual calculation process, theZero-Crossing Window, the area subjected to
calculation shown with dotted lines in Fig. 5, is set,
and the multiple sensor signals included in thiswindow are statistically processed.2) A block diagramof the angle calculator circuit is shown in Fig. 7. The
sensor data AD1 through AD4 obtained from thesensors on the four edges are stowed in the memory.
Then the data within the zero-crossing window aresampled and subjected to calculation. In the calculator
circuit, the positions of the two zero-crossing points
are determined accurately to enable linear
approximation based on a Least Mean Square (LMS)
method. Using the coordinates of the two determinedpoints, calculation is executed as per expression (2)(arctan CORDIC circuit), and then the calculated 10bit angle data is output to an external system. Since
this calculator circuit includes a communication circuit(Comm.I/F), it can output angle data through a serial
communication means.

θ= tan-1 (Δy/Δx)…………………………（2）Fig. 4 Arrangement of sensor arrays(magnet is set at an angleθ)Fig. 7 Block diagram of the angle calculator
Fig. 5 Output signal of the sensor arrays-100-Development of a High Precision Angle Sensor

MAGFET. Use of a native-MAGFET leads to thefollowing advantages.

2.4 Features of our sensor system
Our sensor system calculates an angle based onmagnetic field distribution measured by multiple
magnetic sensors, and has the following advantages.
(1) Since it is intended to detect a position where amagnetic field does not exist (position where thelevel of magnetic sensor output is zero), our sensor
system is not affected by sensitivity variation or the
non-linearity of the sensor elements. As a result, acomplicated compensation process is notnecessary.
(2) Since multiple sensor signals are statisticallyprocessed, the effect of signal noise is minimized,
resulting in the improved angle detection accuracy.
(3) Since the angle is calculated from the coordinatesof two zero-crossing points, our sensor system is
not much affected by the positional difference
between the magnet and the sensor elements.
(4) The MAGFET magnetic sensor chip can be
manufactured by a common CMOS process. As aresult, related circuits can be incorporated into thesame chip, reducing manufacturing costs.

(1) Compared to the p-well region, the impurity
concentration on the p-substrate is lower. This inturn leads to higher electron mobility within thenative-MAGFET, causing the MAGFET chip tohave higher magnetic sensitivity.
(2) Because the p-well region is formed through an ion
implantation process, the concentration of
implanted ions can vary, which leads to variation in
the characteristics of the sensor elements formedin this region. Native-MAGFETs are free from thisproblem, so variation of characteristics is
minimized.
The results of measured magnetic sensitivities ofnormal and native MAGFETs are plotted in Fig. 9. In
this diagram, IMAGFET represents bias currents
carried by the MAGFET. While the magnetic sensitivityof the normal-MAGFET was 2.7%/T, that of the nativeMAGFET was approximately 5%/T%. The sensitivityof the native-MAGFET is approximately twice as highas that of the normal-MAGFET. At the same time, welearned that the magnetic sensitivity of the nativeMAGFET decreased as the bias current increased.
Because the sensor elements were formed directly on
the p-substrate, the channel structure of regions withlower gate voltage seems to be similar to that ofembedded channels. We believe that with a higher
gate voltage, the distribution of channels shifts to thesurface. Since the channels are affected by diffusion
at the interface, decreasing electron mobility,
magnetic sensitivity is reduced.

3. Native-MAGFETAs explained above, our MAGFET magnetic sensor
chip can be manufactured by an ordinary CMOS
process. However, because it is made from a siliconmaterial, its magnetic sensitivity is lower than Hallsensors and other sensor chips made from chemicalcompound semiconductor materials. In addition,
variations in the characteristics of the MAGFET are
unavoidable because the chip is manufactured
through an ordinary CMOS process.
To attain sufficiently accurate angle detection with
our sensor system, it was necessary to improve thesensor performance by enhancing the magnetic
sensitivity of the sensor elements. To achieve this, weinvented a native-MAGFET structured as shown in
Fig. 8.3) This diagram shows cross-sectional views of
CMOS circuit transistors and the native-MAGFET.
Structured identically to an nMOS transistor, a"normal-MAGFET" is usually formed on the p-wellregion. In contrast, native-MAGFET is formed directlyon the p-substrate, and its impurity concentration
directly beneath the gate differs from that of a normal-

Fig. 9 Magnetic sensitivity of the MAGFETs(IMAGFET : bias current in a MAGFET)

Fig. 8 Cross section of native-MAGFET-101-

NTN TECHNICAL REVIEW No. 73（2005）

4. Verification of our prototype chip
operation

5. Sensor bearing sample
A sample sensor bearing that incorporates oursensor chip is shown in Photo 2. Its structure isshown in Fig. 11, together with the structure of aconventional sensor bearing.

The prototype sensor chip is shown in Photo 1.
This chip was fabricated through a standard 0.25μm
CMOS process [5-layer metal wiring, 1-layer
polysilicon, MiM (Meal-Insulator-Metal) capacitor]. The
sensor chip measures 5 mm by 5 mm, and the sensor
array measures approximately 4.2 mm square.
The technical data of the prototype sensor chip are
as follows.
(a) Number of sensor elements: 184 pixels×4 lines(b) Structure of one pixel: six MAGFETs connected inparallel(c) Chip size: 5×5 mm(d) Chip thickness: approx. 300μm(e) Number of IO terminals: 100 (including test
terminals)(f) Angle detection resolution: ±0.36 degrees

Sensor

The result of angle detection with our sensor chip issummarized in Fig. 10. The horizontal axis
corresponds with one revolution of the magnet (360
degrees). The angle resolution was within ±0.36
degrees, and accuracy of 10 bits per revolution wasachieved.

Photo 2 Picture of a sensor bearing sample

Hall sensor

Sensor chip

SHA-ADCSHA-ADC

Angle
Calculator

MAGFET Array

SHA-ADCSHA-ADCMAGFET Array

MAGFET Array

Multi-pole ring(a) Conventional
sensor bearing

Permanent magnet(b) Newly developed
sensor bearin

Fig. 11 Cross section of sensor bearings

MAGFET Array

Photo 1 Micrograph of a prototype chipFig. 10 Angle detection result-102-Development of a High Precision Angle Sensor

The optimal resolution with conventional sensor
bearings was approximately 2 degrees (180pulses/revolution). As pulse output types, they cannot
detect absolute angles. In contrast, our sensor bearing
is capable of 0.36-degree resolution and can detect anabsolute angle and output this angle as a digital value.
The sensor chip can be installed easily, and thesensor bearing can be of a more compact design. The
angle detection accuracy of the new sensor bearing isplotted in Fig. 12. The detection error at a sampling
speed of 10 kHz was within ±0.3 degrees.

6. Conclusion

Detection error [degrees]

We have developed a novel rotational angle sensor
that employs magnetic sensor arrays. To improve thecharacteristics of sensor elements on silicon
substrates, we developed a new sensor structureusing a native-substrate MAGFET to enhance themagnetic sensitivity of the sensor elements, thereby
achieving a very high angle detection accuracy of 10
bits per revolution (0.36 degrees). This degree ofangle detection accuracy was also achieved with asensor bearing that incorporates a MAGFET sensor
chip sealed in a compact package.
Automotive components are becoming increasingly
more compact and functional. We intend to apply oursimple, compact sensor chip to automotive
applications including steering systems andaccelerator pedals.

References
1）S. Kawahito, T. Takahashi, Y. Nagano, and K.
Nakano, "CMOS Rotary Encoder System Based on
Magnetic Pattern Analysis with a Resolution of 10b
per Rotation," ISSCC2005, Dig. Tech. Papers, IEEE
Int. Solid-State Circuits Conf., pp.240-241, 2005.
2）K. Nakano, T. Takahashi, S. Kawahito, "Angle
Detection Methods for a CMOS Smart Rotary
Encoder," Journal of Robotics and Mechatronics
Vol.17(4), pp.469-474, 2005.
3）Toru Takahashi, Kazuhiro Nakano, and Shoji
Kawahito, "CMOS Magnetic Sensor Arrays Using
Native-Substrate MAGFETs for a Smart Rotary
Encoder," Proc. IEEE SENSORS 2004, pp.973-976,
2004.

Magnet rotational angle [degrees]Fig. 12 Angle detection result using a sensor bearing

Photos of authors

Toru TAKAHASHI

Yoshitaka NAGANO

Shoji KAWAHITO

New Product Development
R&D Center
Mechatronics Research Dept.

New Product Development
R&D Center
Mechatronics Research Dept.

Shizuoka University
Research Institute of Electronics
Professor-103-

NTN TECHNICAL REVIEW No.73（2005）[ New Product ]Tapered Roller Hub Bearings for Large Truck

Hiroshi KAWAMURA**
Akira FUJIMURA**

Hub bearings for larger commercial vehicles have advanced to unit type for the purpose of
reliability and easy assembling. This article features the transition of continuous improvement in hub
bearings for larger commercial vehicles and technology of the newly developed Tapered Roller Hub
Bearing, that has supposed the largest diameter in the world.

1. Introduction

2. Construction and features of large
truck axles

Recently, unit designs have been increasingly
implemented in front axles for large trucks in order toimprove reliability and assembly ease. Previously, two
single-row tapered roller bearings had been used forthis purpose. However, this arrangement occasionallyposed safety problems. Incorrect adjustment of
preload on an assembled bearing during truck repair
caused premature flaking or seizure. To address thisproblem and improve truck safety, unit designs foraxle bearings have been increasingly introduced.
Following European truck manufacturers, Japanesetruck manufacturers also began developing unit
design axles, and now some Japanese truck
manufacturers are using 2nd generation tapered roller
hub bearings (hereafter, GEN2 HUR). This paper
describes the current trend in axle bearings on large
trucks as well as the GEN2 HUR that NTN has
recently developed. This unit consists of an outer race
integrated with a hub ring and has the largest flange
diameter in the world.

Previously, two single-row tapered roller bearings
were usually used for large truck axles. Then, GEN1
units composed of two bearings and set light bearings
were adopted. Compared to previous designs that use
two single-row tapered roller bearings, these designs
boast higher reliability and ease-of-assembly. In 2000,
NTN began mass-production of GEN2 HUR for the
first time in Japan. This bearing type includes a unit
comprised of a bearing and a hub ring mounting
flange (a bearing accessory). Table 1 summarizes thehistory of HUR development by NTN, and the featuresof HURs of various generations. We believe that thetrend for large truck axle bearings will be a shift toGEN2 HURs. To satisfy this need, NTN has
developed a GEN2 HUR that features the world's
largest flange diameter. Furthermore, to attain higher
reliability, NTN is currently developing GEN3 HURs
that do not need axle nut tightening control.

*Automotive Sales Headquarters Automotive Engineering Dept.

-104-

Tapered Roller Hub Bearings for Large Truck

Table 1 Features of Tapered roller hub bearings in each generation
Ranking：☆＜☆☆＜☆☆☆

Non-driven wheel use

Conventional

GEN1

GEN2

GEN3

Rotation system

Outer race rotation

Outer race rotation

Outer race rotation

Inner ring rotation

Assembly and service ease

−

☆

☆☆

☆☆☆

Compactness

−

☆

☆☆

☆☆

Built-in seal

−

☆☆

☆☆

☆☆

Preload control

−

☆

☆☆

☆☆☆

Generation

Driven wheel use

Generation

Conventional

GEN1

GEN2

Rotation system

Outer race rotation

Outer race rotation

Outer race rotation

Assembly and service ease

−

☆

☆☆

Compactness

−

☆

☆☆

Built-in seal

−

☆☆

☆☆

Preload control

−

☆

☆☆

3.2 GEN2 HUR Development

3. NTN HUR development status

NTN has developed a large size GEN2 HUR thatcan be used not only for large truck front axle nondriven wheels, but also for driven wheels on rear full
floating axles. We will describe this newly developed
product in more detail. This GEN2 HUR is alightweight tapered roller hub bearing that has anouter race integrated with a hub ring, and features thelargest GEN2 HUR flange in the world.

3.1 Mass-production of GEN2 HUR
GEN2 HUR technology is described in NTN
Technical Review No. 70. NTN's GEN2 HUR massproduction record is summarized in Table 2.

Table 2 GEN2 HUR Mass production results
Truck
manufacturer

Truck type

Application

Beginning ofmass-production

Company A(Japan)Large truck(24 t)

Front wheel(non-driven wheel)

2000〜

Large truck(24 t)

Front wheel(non-driven wheel)

2001〜

Front wheel
Medium truck(non-driven wheel)(12 t)

2002〜

Trailer rear wheel(non-driven wheel)

2003〜

Company B(Europe)Company C(Europe)

Trailer(1) Bearing structure and assembly
Fig. 1 illustrates the structure of GEN2 HUR for
non-driven wheels, and Fig. 2 shows the structure ofGEN2 HUR for driven wheels. Featuring an outer race
integrated with a hub ring, the newly developed HUR
has been optimally designed to reduce the stress that
occurs on the hub due to high load when the truck
turns as well as to lighten the bearing. Fig. 3 shows
an example of GEN2 HUR application. The assembly-105-NTN TECHNICAL REVIEW No.73（2005）

8.3

method for the GEN2 HUR is the same for both driven
and non-driven wheels.

132
Hub ring(2) Hub ring, inner race and roller materials
The hub ring is made of a unique NTN bearing
material that features good castability and excellent
shock resistance. Areas where greater stress canoccur when the bearing operates are treated with an
induction heating process to give them higher static
strength and rotary bending fatigue strength. The
material for the inner ring and rollers is SUJ2.

φ50

Inner raceCage(3) Lubricant
For axle bearings on large trucks, we are adopting a
newly developed grease that boasts fretting
resistance, high temperature resistance and long life.
Table 3 summarizes the grease's generalcharacteristics, and Table 4 provides the results of ahigh temperature durability life test.
1 The grease includes a base oil that excels in oil film
formation performance at high temperature andunder high bearing pressure in order to prevent
overheating and scuffing-induced wear that tends
to occur on inner race larger ribs.
2 With a urea-based thickener featuring excellent
shear stability, the grease boasts good durability.

Fig. 1 Front non-driven GEN2 HUR structure

75.65

φ60

59.5

φ252

φ252

RollerTable 3 General characteristics of grease
Thickener
Base oil
Operating temperature range
Hue

Fig. 2 Rear driven GEN2 HUR structure

Urea-based thickener
Mineral oil + synthetic oil−30〜＋150˚C
Yellow

Table 4 Test result of grease fatigue

HUR

Grease type

Life test result (hr)Current representative grease
for trucks (Sunlight 2)

46

HUR grease

1618※ NTN test data
Test method = ASTM D3336 modified
Bearing = 6204, amount of prefilled grease = 1.8 g,
test temperature =150˚C
Speed = 10,000 r/min, Fr= 67 N, Fa= 67 N

KnuckleShaft

Nut(4) Cage
The polyamide resin cage used can be easily
assembled into the bearing and has excellent heat
resistance, flexibility and durability. This makes
possible a lighter weight at a lower price. Its shape
has also been optimized to promote ease of assembly
into the bearing.

<Assembly of GEN2 HUR on trucks>
1 The GEN2 HUR is inserted over the shaft.
2 The nut at the shaft end is tightenedwith a predetermined torque.

Fig. 3 Structure of axle assembled GEN2 HUR-106-

Tapered Roller Hub Bearings for Large Truck(5) Seal(6) Hub ring (outer race), inner race raceway
surface, and roller rolling surface and large
end face forms

A double-lip high-pack seal has been adopted toprovide good seal performance against mud water,which trucks traveling long distances often
experience, and to simplify the bearing assembly
process. Table 5 summarizes the results of a mudwater resistance test with various seal types. The sealmaterial for our newly developed bearing is fluorine
rubber (FKM) to ensure better high temperature
durability. The slinger material that provides the sliding
surface of the lip is stainless steel, making it rustresistant.

When a truck turns, a greater load acts on its hub,and an edge load can occur on the hub. To avoid thisproblem, the hub ring (outer race) and inner race
raceway surfaces, as well as the roller rolling surfaces,
have optimal complex crowning (example of contact
stress calculation in Fig. 4).
In addition, the roundness of the large end face ofeach roller has been optimized to prevent scuffing andseizure caused by slippage between it and the innerrace large ribs.

3.3 Test rig and evaluation result
Table 5 Test result of seal performance against mud water
Seal type

Axle bearings are important safety components that
support vehicles so that they can run safely. A benchtest is very important in ensuring the functionality andreliability of bearings. NTN possesses various test rigs
for evaluating axle bearings, and can evaluate HURs
for large trucks through various tests such as rolling
fatigue life test. One of the NTN test rigs and a part of
the HUR evaluation results are described below.

Mud water resistance Durability
Cross-section(durability cycle)
ratio

Conventional seal

2

1

Double-lip seal

4〜7

2〜3.5

(1) Evaluation test rig
High-pack seal

18〜20

9〜10

New seal(double-lip
high-pack seal)

30〜38

15〜19

The NTN durability test machine for large trucks is
schematically illustrated in Fig. 5. This machine iscapable of applying a maximum radial load of 196 kN
and a maximum axial load of 98 kN. In executing
tests, a strain gauge and a measuring instrument areused to measure stresses occurring in various areas.
Since this machine can run with a program that canvary the load and bearing speed during the test run, it
is capable of simulating cyclic loads or actual vehicle
running modes.

※ NTN test data
Speed = 1100 r/min, shaft eccentricity = 0,
as mounted eccentricity = 0.05 TIR,
Liquid level = to the shaft center
Material blocked = Kanto loam powder, JIS class S8, 10 wt%,
Test cycle = 20 h run + 4 h standstill

Contact stress

Outer side inner race
Outer side outer race
Inner side inner race
Inner side outer race

Outer side

Inner side

Fig. 4 Example of contact stress calculation-107-

Evaluationthreshold

NTN TECHNICAL REVIEW No.73（2005）

ELEC. MOTORAXIAL LOADHYDRAURIC CYLINDER

RADIAL LOADHYDRAURIC CYLINDERTEST BEARINGTEST BEARINGTHERMO COUPLEACCELERATIONSENCERTIRE RADIUS

Fig. 5 Test machine structure(2) HUR analysis and testing results

weight reduction of approximately 20%.

The results of analysis and testing of the HUR
illustrated in Fig. 2 are described for three
aspects–hub ring stress analysis, heavy load durability
and high speed durability. In designing the hub ring,wall thickness was thinned through the counterboringof flanges and the optimization of rib shape. For thiseffort, we employed FEM to limit the stresses to levels
below standard values, thereby we have succeeded in

1 Hub ring stress analysis
Predetermined moment loads were applied tovarious areas, a set number of times, to measure thestress occurring in each area. We verified that eacharea successfully withstood load application and did
not reach the fatigue limit. An example of hub ring
stress analysis is provided in Fig. 6.

(Before weight reduction)(After weight reduction)

Fig. 6 FEM analysis example-108-

Tapered Roller Hub Bearings for Large Truck

2 Heavy load durability test (0.6 G turning
moment durability test)

4. Conclusion

The results of this test are summarized in Table 6.
The actual life with each sample was more than threetimes the targeted life (calculated life), showing thateach sample had sufficient durability.

NTN has been mass-producing GEN2 HURs as
axle bearings for large trucks, and has recently
developed a unique GEN2 HUR product that is thelargest in the world. Furthermore, NTN is currently
developing GEN3 HUR products that boast greatly
enhanced reliability. We believe that these HUR
products are excellent axle bearings capable ofsatisfying requirements for long life, improved
reliability and ease-of-assembly. We expect thedemand for them will steadily increase in Japan. NTN
is also mass-producing HURs for driven wheels forrear semi-floating axles on compact commercial
vehicles. We intend to supply various HUR products tosatisfy user needs.

3 High speed durability test
The results of a high speed durability test aresummarized in Table 7. Even if the bearing was
operated at the maximum vehicle speed (160 km/h)
for 20 hours, it did not seize and there was little oilleakage. Thus, we verified that our bearing has
sufficient durability.

Table 6 Test results of 0.6G turning moment durabirity
BearingCalculated Calculated
Number oflife ratio
life
revolutions
4
L（×10 rev） L0（×104rev） L/L0

Front
nondriven
wheel
Rear
driven
wheel

2545
2545

848

State after test

3.0

Suspended

3.0

Suspended

2545

3.0

Suspended

867

3.0

Flaking on innerside inner race

8.2

Suspended

3.0

Suspended

2382

289

867

Table 7 Test results of high speed durabirity
BearingFront
non-driven
wheel
Rear
drivenwheel

Speed（r/min）

326(vehicle speed
160 km/h)

Running
time
Test result（h）State after test

Suspended

n=2
No problems
such as seizure

Suspended

n=2
No problems
such as seizure

20

Photos of authors

Hiroshi KAWAMURA

Akira FUJIMURA

Automotive Sales Headquarters
Automotive Engineering Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

-109-

NTN TECHNICAL REVIEW No.73（2005）[ Technical Article ]The Auto-tensioner Market and Technical Trends

Satoshi KITANO*
Tadahisa TANAKA*
Tomokazu NAKAGAWA*

Automotive engines generally use accessory drive belts, timing belts, or timing chains to transmit
crankshaft rotation to the camshaft and accessories. An auto-tensioner can act to keep the tensionof the belt or chain constant, and has shown to be effective at improving belt and chain life as well as
reducing noise generation.
Although the cost of the system and the number of components increases when auto-tensioners
are implemented, improvements in system reliability and reductions in maintenance are achieved.
Auto-tensioners are required to add constant tension, but also have damping characteristics toallow for rapid tension changes. If the driving methods of the camshaft and accessories are to
improve in the future, the application of an auto-tensioner to increase life and reduce noise will benecessary. Consequently, the design of the NTN auto-tensioner has been improved to help reduce
cost and increase reliability.

1. Introduction

2. Drive systems

To transmit the revolutions of crankshafts to
camshafts and auxiliary mechanisms, ordinary
automotive engines are equipped with engine
components such as timing belts, timing chains andaccessory drive belts. Recently, as automotive engine
performance is being greatly enhanced, therequirements for their components have become
increasingly demanding. In this context, autotensioners play an important role in extending thebelts and chain service lives and in suppressing noise.
This report covers technical trends related toautomotive auto-tensioners.

Examples of automotive engine camshaft andauxiliary shaft drive systems are illustrated in Fig. 1.
Camshaft drive systems for automotive engines canbe categorized into timing belt drive systems andtiming chain drive systems. Both types transmit powerby meshed components. Auxiliary drives systems areusually frictional transmission systems using V-belts.
In all these drive systems, auto-tensioners areplaced in slack spans where the tension is lowest.

*Automotive Sales Headquarters Automotive Engineering Dept.

-110-

The Auto-tensioner Market and Technical Trends

Generator
CamshaftCamshaft

Water pump

Cam sprocket

Idler
Idler
Chain guide

Chain tensioner

Autotensioner

Tensioning
pulleyActuator

Supporting
point

Chain guide
Idler
Direction ofengine shaft
rotation
Air-conditioner

Crank

Direction ofengine shaft
rotation
Crank

Crank

Direction of engine shaft rotation

Auto-tensioner

Timing belt use

Timing chain use

Auxiliary drive belt use (serpentine)

Fig. 1 Layout of Auto-tensioner (example)

share for this drive system type is expanding in Japan
currently.

3. History of camshaft drive systems
The first timing belt drive system in the world wasused commercially in 1946 in the USA, and the first
automotive application there was in 1965. In Japan,
this system was first adopted in 1971. In the 1980's,
this new system began quickly capturing market sharefrom timing chain drive systems.
Through the 1970's, the timing chain drive system
was the main type, but in the 1980's, timing belt
systems became dominant. However, timing chain
drive systems began to be adopted again in the
1990's thanks to the development of high-performancechains. The current market shares for timing chain
drive systems and timing belt drive systems are about
50% each in Japan.
Fig.2 illustrates the direction of camshaft drive
systems.
In the past, all accessory drive systems involved
several belts. In the 1980's, though, serpentine
systems began to be used in the USA. The market

100

4. Importance of auto-tensioners
Unlike conventional fixed-tensioners, when autotensioners are assembled into drive systems they arecapable of maintaining a belt or chain tension at aconstant level. This is done by absorbing variations in
tension, including changes in tension due toexpansion and shrinking from engine temperaturechanges and decreased tension resulting from drive
system aging. Thus, auto-tensioners provide aneffective means for extending the lives of drive belts
and chains and for suppressing noise. Although, whencompared with fixed-tensioner systems, adoption of
auto-tensioners can lead to higher actuator valve
system costs and more complicated structures autotensioners have significant advantages. Drive systems
that use auto-tensioners boast improved reliability andgreatly reduced maintenance needs.

OthersPercentage of vehicles

90
80

Timing chain drive systems

70
60
50
40
30

Timing belt drive systems

20
10
0
2000

2001

2002

2003

2004

2005

Year

Fig. 2 Direction of a cam-shaft drive system

-111-

2006

NTN TECHNICAL REVIEW No.73（2005）

Fig. 4 illustrates various timing belt auto-tensioners.
NTN has been developing various hydraulicdampers including a standard type, a short type that iscompact in the axial direction, and an ultra-compact
type.
The integrated pulley-arm type is a unit design
complete with a hydraulic damper and a tension pulleythat reduces costs and has smaller spacerequirements. The built-in pulley type helps save
space since the hydraulic damper is built into thepulley, and it can replace a fixed tensioner.
A performance comparison of various timing belt
auto-tensioners is shown in Table 1.

5. Auto-tensioner types and structures
5.1 Timing belt auto-tensioners
The actuation systems for timing belt autotensioners include hydraulic, screw, and friction
systems. NTN auto-tensioners employ a hydraulicactuation system that excels in durability and tension
adjustment capability to cope with temperature
variations.
Fig. 3 illustrates the structure of an NTN autotensioner. The cylinder is die-cast aluminum to reduceweight.
Pin
Cylinder(die-cast aluminum)

Retainer ringAir
Top level of oil

RodHydraulic fluid

Return spring

Table 1 Performance comparison (for timing belt)
Type
Element

Oil seal

Wear ring
Plunger

Separator

Unit

Valve sleeve
Steel ball

Manufacturer Function※ Reliability

Hydraulic type

Cost

◎

◎

◎

◎

Ａ

◎

○

○

△

B

○

○

○

△

Screw type

C

△

△

◎

◎

Hydraulic type

NTN

◎

◎

○

○

Viscosity type

D

△

○

○

△

E

△

△

△

○

F

△

△

◎

◎

Friction type

Valve retainer

WeightNTN※Tension adjustment function

Plunger spring

Fig. 3 Sectional view of Auto-tensioner

Tension spring
Hydraulic damper
Miniaturization
Standard type

Miniaturization
Short type

Ultra-compact type

Unit design

Miniaturization

Integration
Separate type

Integrated pulley-arm typeFig. 4 Various Auto-tensioner's for timing belt

-112-

Built-in pulley type

The Auto-tensioner Market and Technical Trends

Auto-tensioner
mounting angle

NTN auto-tensioners employ a hydropneumaticconstruction that is free from oil leakage. However, if
this type of auto-tensioner is mounted at an angle
close to horizontal, the liquid can approach thepressure chamber, possibly letting air enter andleading to damper failure. Therefore, the mounting
angle of this auto-tensioner is limited, and therecommended mounting angle range is ±45˚ fromvertical.
If a greater mounting angle is unavoidable, an autotensioner with a reservoir is available (see Fig. 5).
In the design phase, we create 3D models for thecylinder and pulley bracket of the auto-tensioner tocheck for the possibility of interference. (See Fig. 6.)
Through FEM analysis techniques, we haveachieved both a high level of durability and a compact,
lightweight design. (See Fig. 7.)

Reservoir
Top level of oil

Pressure chamber

Fig. 5 Auto-tensioner with reservoir room

5.2 Timing chain auto-tensioners
Recent camshaft drive systems employ timing
chains rather than timing belts so that the wholeengine length can be shortened and because
reliability is high and running noise is low with timing
chains.
With ordinary timing chain auto-tensioners(hereafter referred to as chain tensioners), engine oil
supplied from the engine is introduced into the chain
tensioner to generate hydraulic damping power. A noback mechanism is included to prevent inward
pressure before engine oil is supplied after the engine
is started. No-back mechanisms can be categorized
as ratchet, serrated screw, ring, and screw types.
NTN uses serrated screw and ring types. To reduceweight, die-cast aluminum cylinders are used. Mass
production of the serrated screw type began in 1999,
and that of the ring type started in 2003.
The constructions of our chain tensioners areillustrated in Fig. 8.

Fig. 6 3D-modelFig. 7 FEM analysis result

Relief valve
Air expulsion screw

Return spring
PlungerScrew rod

Return spring
Relief valve

Spring seat

PlungerCheck valve
Rod seat
Assist spring

CylinderSerrated(die-cast aluminum)
Air expulsion
screw
orifice
Check valve

Cylinder(die-cast aluminum)

Fig. 8 Various Auto-tensioner for timing chain

-113-

Ring

FixtureNTN TECHNICAL REVIEW No.73（2005）

The results of a performance comparison of various
timing chain auto-tensioners are summarized in Table
2.

As shown in Fig. 9, the chain running position
fluctuates greatly (the guide position changes) as
the temperature and running speed vary after theengine starts. The chain tensioner clearly responds
quickly to changes in the chain running position,and presses or pulls accordingly.

Table 2 Performance comparison (for timing chain)
Manufacturer

Type
Serratedscrew type

NTN

Ring type

Basic structure
DamperNo-back
Hydraulic andSerrated screw◎◎
serrated screw(stage-less)

2 Static no-back characteristics

Cost

When the engine is at a standstill, and if the
span where the chain tensioner is present is in atensioned state because of a phase difference
between the camshaft and the valve, then thetensioner will be pressed in. Then, after the engine
is started, and before the engine oil is supplied, theamplitude of vibration could be greater due to theamount of inward pressure, possibly causing
abnormal noise. Fig. 10 summarizes measured
static characteristics of the no-back mechanism.
The results in this chart were obtained byallowing a static load to be applied to the plunger
tip and measuring the resultant plunger
displacement until a particular load was reached.
Though the amount of inward pressure relative tothe screw axial clearance is apparent, no further
inward pressure occurred because the screw
remained locked.

△

Hydraulic

○

Ring

○

◎

Ratchet typeA

Hydraulic○Ratchet○○Ring typeB

Hydraulic

○

Ring

○

○

Ratchet typeC

Hydraulic○Ratchet○○Screw type

D

Screw

△

Screw

△

△

5.2.1 Serrated screw functions
With a no-back mechanism that employs rings,
there are actuation stages, and the return allowance isgreater. In contrast, with serrated screw types, the noback mechanism can function regardless of the screw
position, actuation is stage-less, and the return
allowance is equivalent to the backlash and can be setvery small.
Unlike ordinary screws, serrated screws areconstructed so that they are still when a static load isapplied in the inward pressure direction, but operate
swiftly in the outward direction.
The results of an investigation into the functions ofserrated screws are described below.

Measured three times

0.20

0.20

Screw axial clearance

0.20

1 Results of plunger tracking verification
Load（N）

Any auto-tensioner must be able to quickly
absorb changes in tension due to engine thermal
expansion or due to chain or belt slackening whenthe engine starts.

6500r/min（25˚C）

3000r/min（25˚C）

Guide displacement mm

Outward direction

Standstill (25°C)

5

3000r/min（80˚C）

Plunger position change
4Fig. 10 Static characteristic of 'No-back' mechanism
500r/min（25˚C）

3Start-up

500r/min（80˚C）

RestartStandstill

2
0

5

10

15

20

25

30

35

185

190

Time elapsed min

Fig. 9 Plunger position during engine operation-114-

The Auto-tensioner Market and Technical Trends

3 Dynamic inward pressure characteristics
As shown in this chart, when the chain tensioner
is actuated, the screw is unlocked, and the plunger
is pushed in, preventing the chain from becoming
over-tensioned.

While the engine is running, the chain tensioner
plunger pushes in to prevent chain snaking, but
over-tensioning the chain must also be avoided.
The results of measurements of the inward
pressing on the actuated chain tensioner are
summarized in Fig. 11.

Frequency: 10 Hz

Load kgf

0
10
20
30

Displacement mm

40

80 loading cycles

8 sec

10.00
9.75

Amount of inward pressing after 80 loading cycles：0.150mm

9.50

Fig. 11 Dynamic characteristic of screw plunger

The currently available NTN chain tensioners are forfour-cylinder and six-cylinder engines. With a fourcylinder engine, a rotational 2nd-order varying load isapplied to the chain tensioner, and with a six-cylinder
engine, a rotational 3rd-order varying load is applied.
With a six-cylinder engine, the maximum vibration
frequency of a varying load can reach 300 Hz. Whenthe speed of a four-cylinder engine on a motorcycle
reaches 15,000 r/min, the frequency of the varying
load can reach 500 Hz. Every chain tensioner must be
capable of coping with this high-frequency range.
Fig. 12 illustrates the results of response relative tovibration frequencies applied.
This graph shows that the load is generated stably
to a maximum frequency of 500 Hz, and that the chain
tensioner reliably responds to the varying load.

5.2.2 Damper adjustment mechanism
The chain tensioner receives a varying load fromthe chain via a hydraulic damper. The hydraulicdamper setup is optimized based on the relationshipbetween the chain tension and the guide amplitude(plunger amplitude).

1 Method for adjusting the hydraulic damper
The damping force is determined by the leakclearance and relief valve opening pressure.
The user must check the effect of the leakclearance and relief valve opening pressure inadvance and optimize the tensioner reaction force
settings.

2 About tracking
The varying load acting on the chain tensioner isgoverned by the number of engine cylinders.

Generated load N

2000

Plunger amplitude: ±0.05mm
1500
1000

Maximum level500
Minimum level
0
-500
0

100

200

300

Vibration frequency Hz

Fig. 12 Dynamic-characteristic evaluation result

-115-

400

500

NTN TECHNICAL REVIEW No.73（2005）

The construction of an NTN auto-tensioner isillustrated in Fig. 13.
A performance comparison of various autotensioners for automotive auxiliary mechanisms issummarized in Table 3.
In the design phase, we developed a 3D model forassessing the shapes of pulleys and pulley brackets,
checked for interference with engine-related
components and performed FEM analysis to achieveboth reduced weight and higher rigidity.
In order to help promote environmental
conservation by improving automobile fuel economy,
an idling-stop mechanism for automotive engines isavailable. With this scheme, the integrated starter
generator (ISG) drives the accessory belt to turn the
crank to restart the engine. NTN has been developing
hydraulic type accessory belt auto-tensioners
optimized for ISG engines.
We believe that the demand for hydraulic autotensioners will be increasing because they will enable
longer drive belt life as loads from auxiliary
mechanisms, including larger alternators and greater
air-conditioner capacities, increase.

5.3 Auto-tensioners for accessory drive belts
Previously, commonly employed belt tensioners
used multiple belts with one or two auxiliary
mechanisms driven by each belt. With thisarrangement, it was necessary to adjust the tension ofeach belt while running the auxiliary mechanisms andturning an idler pulley and measuring the tension byobserving the amount of belt sag or by measuring thebelt tension with a tensiometer.
Recently, serpentine designs have frequently beenused to drive all the auxiliary mechanisms with one
belt because of reduced maintenance needs, ease ofbelt installation, reduced weight and shorter total
engine length. At first, serpentine designs were used
for large gasoline engines with six or more cylinders,
but they have been increasingly used for gasoline
engines and diesel engines with four or fewer
cylinders.
However, the longer belts of serpentine systems areassociated with elongation that occurs when driving
the auxiliary mechanisms or as a result of aging.
Therefore, an auto-tensioner must be incorporated toautomatically adjust the belt length and maintain thetension at a constant level. Because belt tension
changes due to variations in the load from auxiliary
mechanisms and in the engine speed, the autotensioner must be able to provide a damping force.
Insufficient auto-tensioner damping power leads togreater belt snaking or belt slip (squeaking).
Friction and hydraulic are the two types ofmechanisms available for generating damping forces.
NTN employs a hydraulic type because of its greater
damping force and higher reliability.

Table 3 Performance comparison (for accessory belt)
Type
Hydraulic
Frictional

Manufacturer Function Reliability Weight

NTN

◎

◎

○

○

Ａ

○

○

○

○

B

△

△

◎

◎

Linkage
Spring seal(die-cast aluminum)
Retainer ring
Oil seal
Rod
Return spring
Check ball
Valve retainer
Plunger spring
LinkageFig. 13 Auto-tensioner for accessory belt

-116-

Cost

Cylinder(die-cast aluminum)Air
Top level of oil
Hydraulic fluid
Wear ring
Plunger
Valve sleeve (steel)

The Auto-tensioner Market and Technical Trends

6. Conclusion
We have introduced the trend for various autotensioners related to improved reliability in driving
camshafts and accessory shafts. We are determined
to develop less expensive, highly reliable autotensioners that feature improved functionality andconstruction.

References
1）Kazuki Kawashima, NTN Technical Review No. 61(1992)
2）Katsumi Furutani, Kazuki Kawashima, NTN Technicalreview No. 65 (1996)
3）Satoshi Kitano, Tadahisa Tanaka, Tomokazu
Nakagawa, Monthly Tribology (2004, 10)

Photos of authors

Satoshi KITANO

Tadahisa TANAKA

Tomokazu NAKAGAWA

Automotive Sales Headquarters
Automotive Engineering Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

Automotive Sales Headquarters
Automotive Engineering Dept.

-117-

NTN TECHNICAL REVIEW No.73（2005）[ New Products Information ]Advanced Technology Deep-Groove Ball Bearings forHigh-Speed Servomotors
Higher speeds are capable with these prelubricateddeep-groove ball bearings!

PrelubricatedDeep Groove Ball Bearings

Features

62XX and 63XX series

Capable of dmn value 95×104¡Optimal design improves lubrication
reliability

Newly-shaped synthetic resin cage
High-speed, long-life grease

Rapid temperature increase test

High-speed, high-temperature durability test
6000

100

New cage
Crown-shaped cage

80

Test duration hOuter race temperature increase ˚C

Maintenance-free: No need to grease or clean, and usable life is three times the length

60
40
20

5000
4000
3000
2000
1000

0
3000

6000

9000

Rotation speed min-1

12000

［

15000
dn=600,000
dmn=950,000

0］New cage+
ME-1 grease

6209 with non-contact seals on both sides
Bearing
description （φ45mm×φ85mm×19mm）
Rotation
Blue：n=11,111min-1，Red：n=13,333min-1
speed
Bearing
130˚C (bearing outer ring)temperature

6308 with non-contact seals on both sides
Bearing
description （φ40mm×φ90mm×23mm）
GreaseSteel plate wave-type cage+
MP-1 greaseME-1 grease common to both cage
specifications-118-

New Products Information

Idling Shudder-Suppressing Fixed
Constant Velocity Joint
We have developed a constant velocity jointthat features the lowest level of idling shudder in the world.

Features

Application on actual vehicle¡Compared to similar competitor products, the
new NTN fixed type constant velocity joint
features a 40% reduction in bending load in
the operating angle direction, and a 20%
reduction in actual vehicle idling shudder.

Improvement by adoption of low slide-resistance CVJ (PTJ)
Differential gearing
Absorption by sliding

Vibration from engine

Absorption
by bending
Shaft

Performance

Fixed joint

Shudder acceleration

Idling shudder (engine speed: 620 rpm)

By means of bending rigidity-reduced CVJ

20% reduction

EBJ-M(new CVJ)

Competitor
product-119-

Sliding joint

NTN TECHNICAL REVIEW No.73（2005）

Low Profile Thrust Needle Roller Bearings
NTN has developed the world's thinnest low profile
thrust needle roller bearing!

Features¡These thrust needle roller bearings with sheet steel cages possess thicknesses as low as 1 mm, thethinnest in the world.
Part design: The roller fixing parts of the cage were optimized to eliminate possible interference
Press-flattening process
for roller fixing parts

Single-sheet
high-strength cage

Ironing process for roller fixing parts(that also serve as roller guides)

Roller guide section

NW type (low profile cage)

FW type (ultra-low profile cage)

Roller diameter: between 1.5 mm and 2.0 mm

Roller diameter: between 1.0 mm and 1.5 mm

W type (conventional cage)●No interference between the roller
fixing parts on the cage and therollers(New)
NW type (low profile cage)●Interference avoided by pressflattening the pockets of the cage
●Though having the same sheet thickness as
the W type cage, this bearing cage features
both low profile and increased strength(New)
FW type (low profile cage)●Interference is avoided by ironing thecage pockets●The single sheet form has a low
profile, as well as higher strength fromincreased thickness

Roller diameter: 2.0 mm

Roller diameter: 1.5 mm

Roller diameter: 1.0 mm

Sheet thickness: 0.4 mm

Sheet thickness: 0.4 mm

Sheet thickness: 0.7 mm(Cage strength)
1.0(Cage strength relative to W type)
1.3¡Reduction of moment load at stressconcentration areas by low profile design(Cage strength relative to W type)
3.0¡Reduction of moment load at stressconcentration areas by increased sheet
thickness and low profile design-120-

New Products Information

S-TITAN™ Series of Self-Aligning Roller Bearings
Sets new standards with longest life and greatest load capacity
in the world!

Features

High temperature life
3000

Test conditions
Test temperature:
200˚C
Test piece:
dia. 47×7 mm
flat plate
Contact stress:
5.5 GPa
Load frequency:
3000 r/min
Lubrication:
Ether oil(oil bath)

Load capacity comparison (dynamic load
rating) with competitor product2500
No. of load (cycle)¡High temperature long-life bearing steel that wasdeveloped with consideration for the global environmenthas been adopted for the standard series of self-aligningroller bearings. The internal design of the bearing employsa center-rib free inner race and symmetrical rollers.
¡Long life in ordinary to high temperatures¡Strong resistance to surface failure¡Better dimensional stability at higher temperatures¡Increased crack fatigue strength¡Simplified replacement inventory management
Optimization of internal design and surface finish ensuresconsistent high performance.

2000
1500

1000
500
0
Current

Ordinary temperature life(%) 160

500

Test conditions
Test bearing:
22208

150

400

(dia. 40×dia. 80×23)

140

Test load:
Dynamic load
rating ratio = 0.5
Rotation speed:2000 min-1
Lubrication:
Turbine oil(oil bath)

130
120
110

Life (h)Comparison of dynamic load ratings(NTN/Company A)

S-TITANTM

300

200

100

100
90

0
0

50

150

250

Bore, mm

-121-

Current

Company A

S-TITANTMNTN TECHNICAL REVIEW No.73（2005）

Bearing with Magnetic Field Resistant Rotary Sensor
Magnetic field resistance: three times that of previous products

Features

Structure¡Sensor can be built into a high-torque motor,
allowing very compact systems¡Sensor rarely malfunctions even in
environments with strong external magnetic
fields¡Built-in sensor allows greater compactness
and easier assembly¡Sensor can detect rotation even at speeds
near 0 rpm¡Sensor output signal can be mathematicallyprocessed, allowing calculation of rotational
angle and velocity and detection of rotational
direction

Magnetic sensor

Magnetic encoder

Performance

Specifications

Resistance to external magnetic fields
4

Supply voltage

5〜24V, DC

3

Ratio

Allowable temperature range −40˚C〜120˚C
Rectangular wave, open-collector
Output types
Applicable bearing numbers

2

Phases A and B

16202〜6209

0

Product on the market-122-

NTN productNew Products Information"Ultra" Compact Torque Limiter
for Office Automation Equipment
×12.
First in the bearing industry! Development of an ultra compact size of dia. 14×
Great space savings in office automation equipment.
Conventional
product

New productUltra compact size
Dia. 14×12 mm(conventional product: dia. 18×18 mm)

Bi-directional rotation capability
Easy paper-jam maintenance

Conventional
product

Conventional
product
RubberrollerNew
product

Rubber roller diameter can be made smaller and the axial
direction more compact (red line shows a conventional
torque limiter).

New
product

NTN's new torque limiter can be built into rubber rollers,
allowing reduction of the number of components (blue areashows NTN''s new torque limiter).

Performance
40

Torque stability is equivalent tothat of conventional products

Torque mN・m

35
30
25
20

<Test conditions>Temperature: Ambient / room temperature

15
10

Operation pattern: Actuation for 1.1 seconds
and no motion for 0.2 seconds
Rotation speed: 220 min-1

New product
Conventional product

5
0
0

50

100

150

200

250

Rubberroller

300

Time h-123-

NTN TECHNICAL REVIEW No.73（2005）

RustGuard™ Highly Corrosion-Resistant Bearing for
Steelmaking Machinery
Special coating improves bearing corrosion-resistance and extends life to
3.5 times that of conventional products.

Features

Application¡Longer bearing life with improved corrosion
resistance¡Use in steelmaking and papermaking
machinery parts that are exposed to waterWith our newly developed coating technique that forms a
special manganese phosphate film, we have succeeded increating a protective film that is two to three times thicker
than that attained with other bearing products. In situations
where corrosion (such as roller mark rusting*) can occur,
RustGuard bearings boast a life 3.5 times longer thanconventional bearings.

Results of failure mode
reproduction test on actual machines*Roller mark rusting: Each roll neck bearing on a rollingmachine is left stationary water present on it during roller
maintenance work. As a result, streak-patterned axial rusting
occurs at equal intervals (roller-to-roller pitches) on the
raceway surfaces on inner and outer races. This phenomenon
is called "roller mark rusting."

Accumulated failure probability %<Conventional product vs. RustGuard>

Roller markrusting

Flaking

Evidence ofbearing failureon actualmachinery

90
80
50
20
10
5

1
106

RustGuard
L10 =7.4×107 cycles
Conventional specification
L10 =2.1×107 cycles
107

108

Life (loading cycles)

-124-

109

New Products Information

Hybrid BEARPHITE Bearings
Hybrid structure of resin and sintered metal achieves low friction,high precision and quiet running!

Features

Structure¡Low friction (μ = 0.05)¡Higher precision than ordinary resin bearings¡Quieter running compared with rolling
bearings¡Capable of carrying an axial load
¡Can be applied to an aluminum shaft

Sintered material:
Copper-based
Resin layer
φ8 φ16Resin
Sintered metal layer

Performance (frictional characteristics
for carrying radial loads)<Test conditions>Radial load: 29.4 N
Rotation speed: 240 rpm
Temperature: Room temperature
Mating component material: SUM + Ni-plating (clearance 15μm)

Dimensions
Bore
diameter

Frictional coefficient μ0.3

Hybrid BEARPHITE
Resin
Sintered bearing
Rolling bearing

0.25
0.2
0.15
0.1
0.05
0

100

200

300

Bearing outside
Length Roller bearing number
diameter

1

φ 6

φ12

4

WBC6-12ZZA

2

φ 8

φ12

3.5

W678ZZA

Wear on resin 19μm

3

φ 8

φ16

5

W688ZZ

Wear on sintered bearing 10μm

4

φ10

φ15

4

W6700LLFV4

Wear on hybrid BEARPHITE 6μm

0

5

400

500

Test time h-125-

NTN TECHNICAL REVIEW No.73（2005）

Mono-Drive Two-Way Feeder™
Breakthrough two-way feeding unit achieves parts storage, aligning andfeeding functions in one linear feeder!

Features

Plan view¡Single linear feeder drives two chutes: one foraligning/feeding and one for returning

Bowl system(conventional)

This novel return feeder is comprised of a newly developed
leaf spring unit installed on a conventional linear feeder togenerate inclined vibration in two dissimilar directions andachieve parts feeding and aligning. (Patent pending)¡Space saving and lightweight
Compared with conventional bowl feeders, this linear aligning
feeder is simple and compact, requiring less than half thefloor space. Since this single linear feeder unit is capable ofstorage, aligning and feeding, it can contribute to overall
equipment weight reduction.

Mono-drive
two-way feeder¡Simple construction makes it the optimum
choice for handling multiple work piece types
and small lot production

Applications

The simple construction allows easier maintenance. The
feeder can handle multiple work piece types by simply
installing chutes that are suitable for the work pieces handled.

This feeder is capable of handling various small or
medium-sized work pieces, including machinecomponents, electronic components and plastic parts.

-126-

New Products Information

Ultra Compact Grinding Sludge Solidification System
Low power consumption, easy operation,and environmental impact alleviation!

Features of ultra compact type

Capacity and applications by type

¡Floor space required (approx. 0.8 m2) is 40%
that of the compact type (approx. 2 m2)¡Size allows for direct connection to amachining line or a grinding machine¡Capable of handling amounts of delivered
sludge as small as 10 kg/h¡Very economical at 50% the cost of, and using
less power than, the compact type

Both types can be used according to needs.

Benefits from adoption of the ultra compact
type (equivalent to conventional systems)

Dimensions and weight of ultra
compact typeType

Handling capacity for liquefied
grinding sludge

Applications
Local processing (easy direct

Ultra※ connection to grinding sludge
compact 10 kg/h (1.8 ton / month)filtration systems)
Centralized processing,Compact 20 kg/h (3.6 ton / month)※ batch processing※These values were obtained by assuming that the equipment is operated
8 hours per day, 22 days per month.

¡Cost reduction by recycling and reuse of oilbased coolant
Grinding liquid recovery ratio of 90% or
higher¡Reduction of industrial waste emissions
Capable of reducing sludge volume to 20%
and weight to 55%¡Reduction of waste transportation and landfill
disposal expenses

Dimensions: W1,000×D800×H1,530 (mm)Weight: 1 ton-127-