special issue; automotive products...ntn technical review no.73 ¢2005 £ < new product > high...

ISSN 0915-0528 OSAKA,JAPAN

October 2005

Special Issue;Automotive Products

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 and technologies 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 the Center seeks to accelerate development of new products and ensure swift technical support for customers around the globe. 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 and recognizing NTN as a valuable engineering partner.

General Research & Development Center Established in an Effort to�Enhance New Product Development Capabilities

[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

Contribution

Our Line of New Products 118

2Technology on Environmental Protection for The Next DecadeNobuo IWAI

1Preface

14Application of Topology Optimization and Shape Optimization for Development of Hub-Bearing LighteningHaruo NAGATANI and Tsuyoshi NIWA

High Capacity Tapered Roller BearingsTakashi TSUJIMOTO and Jiro MOCHIZUKI

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

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

Drawn Cup Needle Roller Bearings for Throttle BodiesHideki AKAMATSU

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Ballscrew for Automated Manual TransmissionKoji TATEISHI and Keisuke KAZUNO

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Tapered Roller Hub Bearings for Large TruckHiroshi KAWAMURA and Akira FUJIMURA

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

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48

76

30

10

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

Dynamic Analysis of a High-Load Capacity Tapered Roller BearingKazuyoshi HARADA and Tomoya SAKAGUCHI

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The Reduction of Hazardous SubstancesKiyoshi NAKANISHI and Masakazu HIRATA

Development of the Mono Ring CVTYuichi ITOH and Tomoya SAKAGUCHI

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

Constant Velocity Steering Joint (CSJ)Kenta YAMAZAKI

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

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

The Auto-tensioner Market and Technical TrendsSatoshi KITANO, Tadahisa TANAKA and Tomokazu NAKAGAWA

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

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Kenji OKADA

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CONTENTS

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NTN TECHNICAL REVIEW No.73（2005）

[ Preface ]

The Kyoto Protocol, adopted in the 3rd session of the Conference of the Parties to the United NationsFramework Convention on Climate Change (COP3) held in Kyoto in 1997, came into effect in February2005, and Japan is now facing its obligation to reduce its emissions of greenhouse gases. Recentautomotive technologies are increasingly focused on the challenges of environmental issues. Asevidenced 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 invitedMr. 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-protectiontechnologies for automobiles. This issue also summarizes new eco-friendly products and eco-conscioustechnologies associated with major NTN products, including constant velocity joints, hub unit bearingsand needle bearings, as well as unit-built and modular products that incorporate electronicstechnologies and other unique automotive bearing products.

Our major products, bearings and constant velocity joints, are essentially eco-friendly products thathelp 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, lowertorque, higher efficiency and longer service life. In developing these new products, we not only considerenergy savings during use by our customers, but also aim to reduce environmental impacts duringproduct 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 ourproducts, to introduce highly efficient production lines, to rationalize logistics systems and to achievezero waste emissions.

In September 2005, NTN established the General Research & Development Center at our IwataWorks to further enhance our product development capabilities in high technology fields including nextgeneration 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 TechnologyNetwork" in promoting its new R&D projects and contributing to society.

Kenji OKADAManaging Director

Introduction to this Special Feature Issue –

NTN's Automotive Products

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[ Contribution ]

Technology on Environmental Protectionfor The Next Decade

1. 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 cartswere driven by people, oxen and horses, while boatsand ships were powered by people and wind. In short,renewable energy resources were used in everyaspect of everyday lives in Japan. In those days,petroleum and coal were available in very limitedareas in small quantities. For nearly three centuriesbefore industrialization and westernization, Japanmaintained a unique seclusion policy in which thecountry conducted trade with only Holland and China.During this period, industrial resources and rawmaterials were rarely imported into Japan. Peoplecontinued to live in a closed land of a limited area onrenewable resources.

In the average farming village, eldest sons inheritedland from their parents. If land had been passed onnot 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

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 alsoto his younger sons was called "tawake", a foolishperson who splits up his paddy fields. This wisetradition of our ancestors was intended to preservelimited 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 greaterpower, amounts which had not previously beenavailable from human energy alone. Eventually,mobility developed by use of transportation means,typified by automobiles. This revolution in mobility wastriggered by the discovery of an affordable energysource-petroleum. According to statistical data fromthe Petroleum Association of Japan, global annualproduction of this resource reached 413 millionkiloliters in 2004. This great amount of crude oil iscollected from oil fields annually, and a large portion ofit is burned, generating gaseous substances, includinggreenhouse gases and particulate matters that arereleased into the air. A portion of it is converted intosolids in the form of polymers that are recycled and

Nobuo IWAI

Senior Chief Researcher FC・EV Center, Japan Automobile Research Institute

Exhaust emission purification on both gasoline engine vehicles and diesel engine vehicles could besolved by 2010 using exhaust catalysts and low sulfur fuels. Global warming gases and energydiversifications 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 unitconfiguration, traction control four-wheel drive using a traction motor could increase not only energysavings, 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 friendlyvehicles.

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 ofpetroleum resources also has negative impacts.

Current challenges for automotive environmentalprotection technologies are gas emissions reductionand global warming prevention, while mid and long-term 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 ifthe shapes, engines and energy sources of futureautomobiles differ from those of current cars. For thispurpose, we need to address the challengesassociated with environmental protection technologiesfor automobiles.

2. Gas emissions control

Gas emissions control for gasoline-engine cars hasbeen nearly solved as the NOx emission level of low-polluting 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. Fordiesel-engine cars that emit not only gaseoussubstances including NOx but also particulate matters(PM), emissions will be successfully controlled in thenear future through improved combustion and catalysttechnologies. The petroleum industry in Japandecided to reduce the sulfur content in diesel oil to 50ppm or lower by the end of 2004 and 10 ppm or lower

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

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during 2007. According to the new 2005 long-termregulation, the combination of low-sulfur diesel oil andlow-pollution diesel engines will help reduce NOxemissions to 12% and PM emissions to 3% comparedto conventional diesel-engine cars. Furthermore,according to the 8th Report from the CentralEnvironmental Council of Japan, NOx emissions willbe reduced to 5% and PM emissions to 1% by 2009.In other words, virtually all the technologicalchallenges related to automotive gas emissionscontrol will be solved in the first decade of the 21stcentury. Though these gas emissions controlinitiatives are limited to the Japanese auto market,Japanese automakers, having grown into multinationalcompanies, must now take responsibility forconservation of the global environment. TheseJapanese-born environmental protection technologieswill contribute to environmental conservation indeveloping nations where rapid economic growth istaking place. Japanese contributions to internationalsociety in the form of military power and financialsupport are often discussed in Japan. However, Ihope that Japan's unique environmental protectiontechnologies will contribute to the prosperity andwelfare of the world.

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 and


available 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 situationcontinues for a prolonged period. So far, according tothe statistical data, the CIF based import price of crudeoil has been proportional to that of LNG. However, theCIF based import price of crude oil has not beensynchronous with that of coal. Recently, however, thechange in resource and energy consumption patternsresulting from dramatic economic growth in developingnations including China, has resulted in increasedprices for many resources and raw materials. In the2000-2020 time span, the demand for crude oil inChina seems to have nearly doubled due to China'sdramatic economic growth. Currently, about 90% ofcrude 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 MiddleEast.

Another uncertain factor about petroleum supply isthe special situation in America. Unlike the OPECbasket, the crude oil price OPEC uses as a standard

price, the price of West Texas Intermediate (WTI), astandard US market brand that is produced in Texas,has always been about $5 per barrel higher. Recently,the difference between the two prices has beenincreasing. For instance, crude oil spot prices as ofOctober 2004 were $37.54 per barrel for the OPECbasket and $53.24 per barrel for WTI, a difference of$15.70 per barrel.

This is because WTI is a superior light crude oil thatcontains less sulfur content, and because the facilitiesthat produce light oil from heavy oil lack capacity. As aresult, the supply capacity for light oils, includinggasoline, kerosene and diesel oil, is insufficientcompared to their demand.

According to the estimated amounts of crude oildeposits 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 miningyears for the USA, China, and Norway and the UKcombined, oil producing nations that also consumelarge amounts of petroleum, are 11, 14 and 7 years,respectivelyﾐsurprisingly small reserves. These miningyears have been obtained by simply dividing the

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■Estimated amounts of crude oil deposits and mining years worldwide (as of the end of 2004)As of the end of 2004, the proved reserves worldwide amounted to approximately 1.28 trillionbarrels with 49 remaining mining years. 69.3% of the proved reserves were accounted for bythe OPEC nations with 57.1% in the Middle Eastern nations.

Unit: million barrels

37year

27,007（2.1％）

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18,250（1.4％）

12,987（1.0％）

14,600（1.1％）

85,986（6.7％）

115,000（8.0％）

Proved reserves: 885,188 (69.3%)Mining years: 84 years

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

Non-OPEC total

Proved reserves: Among the total amount of oil present in a given petroleumreservoir (original oil-in-place), reserves where production is technically andeconomically feasible are referred to as "recoverable reserves" and are saidto 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."

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 bedepleted in 51 years because there are possibilities for new well explorationsand 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)

World totalproved reserves

1,277,702 (100%)

Mining years: 49 years

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Fig.2 The estimated amount of crude oil deposits and mining years in the world (Home page of Petroleum Association of Japan)

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

4. Emission control for greenhousegases and energy conservationmeasures for automobiles

Emission of greenhouse gases released from anycountry in the world, whether or not the country hasratified the Kyoto Protocol, can affect not onlyneighboring 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 livingenvironments of people, we should be aware thatsuch activities also damage the whole of nature.

So that we can continue utilizing convenientautomobiles to maintain mobility while minimizinggreenhouse gas emissions, we must run engines withlow fuel consumption and use sustainable energysources to achieve our travel and transportationpurposes. The very task of engineers who specializein automobiles is to realize highly efficient engines thatoperate on fuels and energies obtained fromresources with stable supplies in order to achieve thisobjective. More specifically, we must realize the costsand systems that are acceptable to society and bring

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

them into common use in order to contribute to energyconservation and reduction of greenhouse gasemissions.

If we can reduce energy consumption for productionand use (traveling) by 50%, greenhouse gasemissions from automobiles will also be reduced by50% and the mining years for crude oil will bedoubled. 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.

(1) DownsizingUsually, users prefer heavy, expensive, large luxury

cars. If most hours of car use are for commuting, carsoften carry only one passenger. For example, imaginea car that weighs 1.5 tons or more traveling with onlyone passenger, leaving four or more seats vacant.Suppose that user purchases a less expensivelightweight car weighing 1 ton or less, dramaticreductions 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 (tendenciesmay vary depending on age group or sex) and learnedthat 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 intotechnologies that help realize downsizing.


(2) Hybrid cars

Stationary engines for driving power generators,etc. can be used as cogeneration systems that offerexhaust heat and cooling heat for hot water supply. Asa result, such stationary engines boast a total thermalefficiency of approximately 80%. In contrast, themaximum net thermal efficiency of automotive enginesis usually in the range of 35 to 45% since the wasteheat 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, includingidling with a thermal efficiency of 0%, are often usedand 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 withoutdissipating that energy as thermal energy. Thiselectricity is accumulated and reused for accelerationof the car. (2) The engine stops while the car is notrunning. (3) In low load operations when thermalefficiency is low, the car runs on electricity that waspreviously generated and stored during medium andhigh load operations when the engine was running athigher thermal efficiency. If the net thermal efficiencyreaches approximately 30% by these measures, fuelefficiency is doubled.

The description above is the basic concept of hybridcars. To realize this concept, systems that storeenergy 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 electricitystorage 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 batteriesmounted in them are less than those in electric cars.This situation may change if energy storage systemsthat are more efficient and useful than electricitystorage systems are successfully developed.However, electricity storage systems will maintain theirprevalence for the near future.

One of the challenges associated with alleviatingautomobile environmental impacts is reducinggreenhouse gas emissions. In regions where a largerportion of power generation is accounted for bynuclear and hydraulic power stations, which havelower CO2 emissions, electric cars are more eco-friendly than hybrid cars. However, the drawback of

electric cars is that the distance they can travel on onecharge is limited, a problem that prevents them frombeing used more widely. In Europe, to overcome thisdrawback, a "plug-in hybrid" system is beingconsidered. These plug-in hybrid cars would functionas externally charged electric cars in urban areas andas internal combustion cars in suburban areas. Incertain situation, they would also operate in hybridmode.

There is also the possibility that a novel type ofhybrid car that features a new concept for reducinggreenhouse gas emissions could be developed.

Having two prime movers, that is, an engine and amotor, hybrid cars can use the driving force of bothwhen a larger output is needed, such as whenaccelerating or climbing a slope. As a result, enginesthat are more compact and powerful can be designedto meet fuel efficiency performance requirements. Forexample, the Honda Civic gasoline passenger car hasan engine capacity of 1688 cc and maximum torque of155 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 maximumsystem torque equivalent to that of the conventionalCivic. 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 systemoutput to as much as 200 kW. Though being SUVs,these cars feature both the low fuel consumptionassociated with popularly priced compact cars and thehigher power of sports cars.

The Toyota Lexus RX300, the Highlander and theEstima Hybrid employ electrically driven 4WDsystems. The front wheels are driven by the engineand the front motor, while the rear wheels arepowered by the rear motor in response to the currentrunning state. In the normal state, these cars aredriven 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 initiatedriving power and the cars switch to 4WD mode.These cars are good examples of how low fuel-consumption power train technology can be fullyutilized to ensure safe driving.

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5. In pursuit of sustainable fuelenergies

Another solution to these social needs of energyconservation and greenhouse gas emissionsreduction that also uses sustainable fuel energies isthe hydrogen fuel cell vehicle (FCV).

In January 2002, the US Department of Energy(DOE) announced the FreedomCAR (CooperativeAutomotive Research) program, which is centered onthe engineering development of hydrogen fuel cells,as the program to succeed the Partnership for a NewGeneration of Vehicles (PNGV). In February 2003,President Bush announced the Hydrogen FuelInitiative 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 ofthe 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 anyUS 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 perkilogram, and a fuel cell cost of $30 per kilowatt powerunit. In 2015, the DOE will make a decision onwhether or not to commercialize the FCV, based onthe 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 FuelCell Partnership (CaFCP) in the USA, the CleanUrban Transport for Europe (CUTE) FC busdemonstration program, and the Ecological CityTransport System (ECTOS) in Europe, and the JapanHydrogen & Fuel Cell Demonstration Project (JHFC).

The FCV commercialization scenario of theJapanese Ministry of Economy, Trade and Industrydefines that the period up to the end of 2005 was forpreparation of engineering fundamentals,demonstration of engineering aspects, establishingthe engineering development strategy, providing legalframeworks, performing demonstration tests anddeveloping fuel quality standards. In the introductoryphase from 2005 - 2010, the fuel supply system will beestablished, 50,000 FCV units will be introduced topublic organizations and enterprises, and the second-stage FC technology development strategy will bedeveloped. In the 2010 - 2020 full commercializationphase, the fuel supply system will be further enhancedand 5,000,000 FCV units will be sold to ordinary users

in order to decrease the price of FCVs. To fulfill thesetargets, government agencies and privatecorporations are jointly continuing research anddevelopment activities.

As has been demonstrated in the JHFC project andelsewhere in Japan, FCVs already have performancesufficient for running in urban streets. Freezing ofwater at cold temperatures is reported to have beensolved by decreasing thermal capacity with a metalseparator 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 membranecapable 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 demandingchallenge for full commercialization of FCVs isachieving a two-digit reduction in price. To realize thisdramatic price reduction, we need to return to the verybasics for research into stack components, such assolid polymer membranes, catalysts and separatorswhere dramatic breakthroughs are expected. Anotherchallenge remaining to be solved is increasing therange on one charge of hydrogen. There have beenresearch efforts and deregulation attempts to increasethe permissible fill pressure of hydrogen containersfrom 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, high-pressure methane hydride stored in high-pressurecontainers appears to be promising.

For some years, the possible methods for producinghydrogen will be onsite production at hydrogen supplystations 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. Inthe future, solar heat, a renewable energy, will beused for steam reforming coal and natural gas toproduce hydrogen without releasing CO2. Japan hasvirtually no useable renewable energy sources,including solar energy. In the future, nations andregions situated in the sunbelt where abundant solarenergy is available, such as the USA, Australia, theMiddle East and Africa, which are also blessed withcoal and natural gas, could become hydrogen supplysources. Methanol and dimethyl ether seem promisingas carriers for the maritime transportation of hydrogen.

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

6. Conclusion

During the next 10 years, hybrid cars will be morecommonly used and fuel cell cars will be increasinglyintroduced. Fig. 4 summarizes the comparison of the10-15 mode fuel consumption of gasoline vehicles(catalog data), hybrid vehicles (catalog data), andhydrogen fuel cell vehicles (actual measurementsobtained from the JHFC project converted to gasolineequivalents). In each vehicle group, the vehicleweights have been adjusted to a common value and theaverage 10-15 mode fuel consumption values and toprunner values are indicated. Since there is variation inthe data of actual cars, comparing the average fuelconsumption values and the top runner values (bestvalues available with the present-day technologies) ofall the groups will be useful in order to evaluate the

Results of 10-15 mode fuel consumption measurements

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

Gasoline density : 0.729kg/LGasoline energy (LHV) : 45.1MJ/kgHydrogen energy (LHV) : 120MJ/kgAverage weight of FCV : 1717kg

10.2 11.1

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Catalog datafor ICV group

Catalog datafor HEV group

Actual measurementsfor 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 theICV 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, Insight FCV 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 Institute

Fig.4 Japanese 10-15 mode fuel consumption on gasoline, hybrid and fuel cell vehicles (Home page of JHFC)

fuel efficiency improvements in each group.As this chart shows, in terms of both the average

and top runner fuel efficiency values, the hybridvehicles are better than the gasoline vehicles, and thehydrogen fuel cell vehicles are superior to the hybridvehicles. Note, however, that this data only shows thedegree of efficiency in the particular period rangingfrom 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. Theauthor believes that data on ever-diversifying fuelproduction techniques will be thoroughly reviewed andautomotive engines will be optimized for the fuel typesused.

In a certain sense, fuel cell vehicles are serieshybrid vehicles whose engines and generators arereplaced with fuel cells. Many components, includingsecondary batteries and motors, are commonly usedin both fuel cell vehicles and hybrid vehicles. If higher

-9-

energy density in energy storage media includingbatteries and capacitors is achieved, for example, orother higher functionality is realized, the number ofbatteries mounted on cars can be decreased, allowinglighter-weight vehicle design.

Furthermore, cost reduction, through improvementin manufacturing techniques for batteries and othercomponents, is key to realizing greater use of fuel cellvehicles.

One of the technologies that helped create hybridvehicles is a unique packaging technology in whichpreviously independent power units, including anengine, motors, and a transmission are built intosingle lightweight assembly that is readily mounted ona car. The Toyota Hybrid System (THS) in which theelectric CVT is configured using planetary gearing is agood example of overcoming various engineeringchallenges.

Currently, an increasing number of models of hybridcars are being introduced into the car market. Someanalysts predict that more than one million hybrid carswill be produced annually. This technical field naturallyinvolves bearing technology, which is a strongpoint ofNTN Corporation. The component technology forhybrid cars will be increasingly used for motive powerin electric 4WD vehicles, electrically drivenwheelchairs and construction machinery. NTN'sunique component technologies, including constantvelocity joints, are vital technologies for powertransmission systems on machines includingautomobiles. The efforts of the NTN staff to improvetheir technologies to realize more compact,lightweight, low-loss products support environmentalprotection technologies and are reflected in bothconventional and hybrid cars. Examples of suchefforts include their E series high-efficiency constantvelocity joints, PTJ low-vibration constant velocityjoints and mechanical clutch units (MCU).


-10-


[ Technical Article ]

The Reduction of Hazardous Substances

1. Introduction

Believing that coexistence with the globalenvironment is our most important challenge, NTNundertakes eco-conscious management aimed at thereduction of environmental impacts and the realizationof a recycling-oriented society. With these goals, NTNis striving to reduce CO2 emissions, the prevention ofair, water and soil pollution, the reduction of wastes,and the promotion of green procurement.

This report describes current and future NTNefforts to reduce hazardous substances in itsproducts.

2. Trend in elimination of the use ofspecific hazardous substances

In 2001, electronic equipment built in Europecontained high amounts of cadmium, a hazardoussubstance exceeding regulated level standards. Thistriggered increased consciousness about greenprocurement.

Consequently, EU issued a series of directives thatcontrol or ban the use of hazardous substances inthese and other products. These include the ELV (Endof Life Vehicle) directive that totally bans the heavymetals lead, mercury, cadmium and hexavalentchromium in cars. The RoHS (Restriction of the use ofcertain Hazardous Substances in electrical andelectronics equipment) directive also imposes a totalban on six hazardous substances, including the heavymetals above and fire retardants for plastics (PBB'sand PBDE's), in electrical and electronics equipment,beginning in July, 2006.

In addition, some manufacturers have alreadystarted various self-imposed efforts to ban certainorganic substances that could harm humans.

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

The NTN Group has continued to achieve environmental harmony according to its EnvironmentalPolicy 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 hazardousmaterials in NTN products. These items would include rolling bearings and constant velocity joints, aswell as many other NTN products.

Kiyoshi NAKANISHI**Masakazu HIRATA**


-11-

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

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

Lead naphthenate, lead dithiophosphate and leaddithiocarbamate are excellent extreme pressureagents that have long been used as additives in high-load greases.

Constant velocity joints, one of our major products,are used in high surface pressure and rolling-slidingconditions. Therefore, greases containing lead-basedextreme pressure additives have often been used onthem to prevent wear and achieve low friction. Incooperation with a grease manufacturer, we havedeveloped alternative greases to address thisenvironmental challenge1), and started to offeralternative lead-free greases to our customers in2002.

Our newly developed greases have uniquecombinations of additives that enhance bearingperformance and contain organic molybdenum,molybdenum bisulfide and similar substances insteadof lead-based extreme pressure additives.

3. Elimination of controlled substances

3.1 Hexavalent chromium: total ban of use inshields and seal cores

Previously, hexavalent chromium had often beenused in NTN product components, including shieldsand rubber seal cores on bearings.

To prevent outward leakage of grease and inwardingress of foreign matter, a shield and a core metalrubber seal are installed on the side faces of sealedrolling bearings. Conventionally, the metal used inshields and cores has been galvanized sheet steeland the surfaces of the pieces have been chromate-treated with a material containing hexavalentchromium.

In 2002, NTN introduced a safe tin-plated sheetsteel 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 electrolytictreatment layer and galvanized sheet steel with aphosphate film.

Photo 1 shows examples of tin-plated sheet steeland conventional chromate-treated (hexavalentchromium) galvanized sheet steel after undergoing a

(Improved) Tin plating(Conventional) Galvanization + chromating

24 hourslater

Rust occurrence

Rust occurrence

48 hourslater

5/5 parts

5/5 parts

0/5 parts

5/5 parts

<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


-12-

Performance data for greases developed for theslidable joint TJ series are summarized in Fig. 1 andTable 1. The results for the 3rd order component ofinduced cyclic axial load that serves as an index forNVH (Noise, Vibration, Harshness) performanceaffecting automobile riding comfort, as well as fordurability, are better than those of a conventionalgrease that contains a lead-based extreme pressureadditive.

In addition, the bearings used in low-speed, high-load conditions, such as for steel-making andconstruction machinery, had employed lithium-mineraloil based greases containing lead-based extremepressure additives. We have already replaced theseconventional greases with lead-free greases that wedeveloped in cooperation with a grease manufacturer.

00 100 200 300 400

20

40

60

80

100

120

Torque (Nm)

3rd

orde

r co

mpo

nent

of

indu

ced

cycl

ic a

xial

load

N

(rm

s)

Conventional grease

Newly developed grease A

Newly developed grease B

Fig.1 The 3rd order component of induced cyclic axial load of TJ

Nitric acidSecondary amine Nitrosamine

R1

NH N NNO2

R2

R1

R2

+ -- =0 + H2O

(In oxidizing condition)

H＋

Fig.2 Reaction of secondary amine and nitrite

Table 1 Light load durability of TJ

Running time hT.P

90 125 300

No.1

No.2

No.1

No.2

No.1

No.2

Type of grease

Newly developedgrease A

Newly developedgrease B

Conventionalgrease

Target hours

○

○

○

○

○

○

○

○

△ △

△ △

○

○

○

○

×

△

(○: No problem△: Minor problem×: Severe problem) Table 2 Composition and properties of sodium nitrite free grease

Conventionalgrease

Newly developedgrease

72

Ether + PAO

Diurea

300

2100

＞300（n=5）

1

100

Ether

Diurea

300

3800

＞300（n=11）

1

Resistance against brittleness-induced peeling h

Life h (at 150°C)

ConsistencyThickener

Base oil viscosity mm2/s

Base oil

Rust-preventive performance(rust occurrence %)

3.3 Cadmium: ban of use in resin colorantA yellow or red cadmium-based pigment has been

used often as a colorant for resins. NTN previouslyused a cadmium-based pigment in certain resin slidingmaterial BEAREE products and in the band resin forexpansion-compensating bearings. The cadmium-based pigment for the former application was replacedwith an alternative material by 1993, and for the latterapplication in 2002.

4. Efforts to eliminate possible toxicsubstances

4.1 Sodium nitrite: ban of use in greasesSodium nitrite has been used often as a rust-

preventive agent because of its excellent performancein this role. It also has high performance as anantiseptic and it is a food additive approved by the USFood 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-solublegrinding coolant because grinding machine operatorsfrequently 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 rust-preventive agents and oxidation inhibitors, and therehave 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 outsourcedgreases 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 hydrogenbrittleness caused by the ingress of hydrogen to thebearings due to the decomposition of the base oils ingreases.

-13-

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

4.2 Phthalates: ban of use in boots and rubberseals for constant velocity joints

Phthalates are often used as plasticizers for plasticsand rubbers. In 2003, certain phthalates were addedto the EU list of regulated substances (EU directive2003/36/EC) as CMR substances (Substancesclassified as Carcinogens, Mutagens or substancestoxic to Reproduction). More specifically, sales ofdibutyl 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 velocityjoint boots contain phthalates as plasticizers. We arereplacing these with phthalate-free components incooperation with associated componentmanufacturers with the goal of eliminating phthalatesby the end of 2007.

*Sales of products containing these substances togeneral consumers are permitted.

4.3 Water-soluble barium compoundsWater-soluble barium compounds are listed as

Class 1 designated chemical substances in thePollutant Release and Transfer Register (PRTR) Law.

Kiyoshi NAKANISHI

EnvironmentalManagement Dept.

Photos of authors

Masakazu HIRATA

Technical Research Dept.Research & Development Center

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 water-insoluble barium sulfonate can generate water-solublesalts. Therefore, we are going to introduce rust-preventive oils and newly developed greases that donot contain barium sulfonate.

5. Conclusion

As stated at the beginning of this report, NTN makescoexistence with in the global environment a toppriority and implements eco-conscious managementaimed at reducing environmental impacts andrealizing a recycling-oriented society. In this report,the authors have introduced NTN's proactive efforts toreduce the use of regulated substances.

Though not described in this report, NTN wishes tocontribute to reducing global environmental impactsand is fully committed to developing eco-friendlyproducts that boast smaller sizes, lower torques,higher efficiency and longer life. Bearings areessentially 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-consciousmanagement activities.

References1) Shinichi Takabe: NTN Technical Review, 68 (2000), 44


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

1. Introduction

The enactment of the US CAFE (Corporate AverageFuel Economy) regulations in 1975 triggeredincreased demand for lightweight designs for everyautomotive component. Lighter components lead tobetter automobile fuel efficiency, which in turncontributes to energy conservation and globalenvironmental preservation. In manufacturingindustries, various efforts to achieve lightweightdesigns are in progress and bearing manufacturersare committed to decreasing the weights of hubbearings.1) 2)

NTN has also been working to lighten variousproducts. By applying shape optimization techniquesto the development of lightweight 3rd generation hubbearings, we have succeeded in dramatic weightreductions. This report describes our optimizationtechniques as well as examples of lightening effects.

2. Development goal

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

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

*Automotive Sales Headquarters Automotive Engineering Dept.

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 topologyoptimization; the second is shape optimization. We introduce the process of the developmentby explaining the two optimization methods shown above.

Haruo NAGATANI*Tsuyoshi NIWA*

Hub shaft

Outer ring

Inner ring

Fig.1 Hub-Bearing original design

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

3. Overview of shape optimization

The basic concept of shape optimization design isto place material in areas that truly need it and thin outunnecessary 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 needsmentioned above, recently the demand for the abilityto determine optimal shapes easily has beenmounting. However, currently available softwarepackages 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 bedescribed in detail later). For topology optimization, anoptimal structure for carrying an external load isdetermined, while for shape optimization, the areas tobe altered and the ranges for possible dimensions areset 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 adjustmentswere made as necessary to achieve optimaldimensions that do not greatly differ from those of thatbasic shape. The reasons for this two-step schemeare as follows.

Necessity of two-step analysis scheme(1) Topology optimization analysis is essentially linear

analysis,* and does not offer sufficiently accurateresults for a boundary nonlinear problem such asstress analysis for hub bearings where contacts

-15-

Flange: design region

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

Shaft: non-design region

Hub ring

Inner side view Outer side view

Inner ring

Wheel spigot jointoutside diameter: non-design region

Bolt-fastening section: non-design region

Fig.2 Topology optimization model

between a flange and a brake rotor or between aninner 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 topologyoptimization are complicated 3D forms that maynot be obtainable by machining at reasonablecosts 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 againsubjected to shape optimization to realize the finaloptimal shape.

*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 optimalstructure 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 optimizationtechnique.

Analysis modelThe analysis model we used is illustrated in Fig. 2.

The hub ring was integrated with the inner ring, meshwas plotted and the 3 translational degrees of freedomfor 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


-16-

Fig.3 Result of topology optimization

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 regionswhere the shape cannot be changed (such as thewheel spigot joint outside diameter and bearingoutside diameter) due to interactions with counterpartmembers were defined as unalterable regions (non-design regions).

Furthermore, analysis was executed considering theforging draft, cyclic symmetry conditions of boltholesand symmetricalness relative to the straight linesconnecting the boltholes and the flange center. Theobject function and restricting conditions were asfollows.

Objective function: Distortion energy minimizationRestricting conditions: 22.5%, 27.5% and 30% relative

to the volume (or mass) of initial shapeThe analysis was performed with two hub bolt

patterns - 0˚ (cross-pattern) and 45˚ (X-pattern)relative to the vertical direction. Since the stress wasgreater in the 0˚ direction, further analysis was limitedto this direction.

30% the initial volume

27.5% the initial volume 22.5% the initial volume

Max

imum

str

ess

MP

a

Str

ess

limit

valu

e

Proportion of initial volume

0.1 0.2 0.3 0.4

Fig.4 Principal stress of each design

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

Wheel spigot joint outside diameter(non-design region)

Analysis resultsThe shapes and principal stresses shown in Figs. 3

and 4 were attained. As shown in Fig. 4, with greaterweight reduction, stress rapidly increased. Weadopted a shape that was 27.5% the initial volumeand near the stress limit as the basis for the basicstructure of the hub bearing.

-17-

5. Shape Optimization (2nd Step)

Analysis modelUsing the shape obtained from the previously

described topology analysis as a basis, the basicshape was set as in Fig. 5 and shape optimizationanalysis was executed by varying the dimensions atthe seven locations (h, b, y1, x2, y2, x3, y3) shown inFig. 6. The fully automatic analysis systemsummarized in Fig. 7 was developed and the analysiswas executed based on the L27 experiment planmethod.

Objective function: Volume (mass) minimizationRestricting conditions: Principle stress at three

designated nodes (yellow dots in Fig. 6)below the current level

The software of the shape optimization analysissystem drives the nonlinear analysis solver to

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

Y

Z X

Height: h Width: b

Fixation at end point of tangential line

X4=const.h, x2, x3<X4

Width: b

Height

（h,y1,b）

（h,y1,0）（x2,y2,b）

（x2,y2,0）

（x3,y3,0）

（x4,y4,0）

（x3,y3,b）

Fixed

y1<Y1where Y1 is currentvalue of y1

Fig. 6 Parameter of DOE

Morphing software

Nonlinear analysissoftware

Execution ofnonlinear analysisShape base vector setup

Design parameter setup

Objective function: analysis result

Optimal solution

Optimizationsoftware

Analysis result

FEM model

Driveoperation

WriteoperationRead

operation

Restrictingconditions

Fig. 7 Shape optimization systems

calculate stress values while altering the shapeaccording to the mesh data.

In altering the shape, the shape base vectors* wereset using the morphing software so that alteration ofthe mesh could be synchronized with alteration of theshape 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 GHzPentium 4 computer and took 2.5 days to complete.

From this result, a response curved surfaceapproximation model was developed to achieve anoptimal solution.

*Vectors that defines how each node shifts whenthe nodes used as parameters are shifted.


Analysis resultsIn Fig. 8, (a) summarizes volumes, while (b) shows

stress values. "L27 best solution" represents thecombination that performed most effectively in thecalculations performed according to the experimentalmethod while "Optimization" represents the optimalsolution obtained by calculation in accordance with theresponse curved surface approximation. As a result ofthis series of calculation operations, many of theoptimal values obtained coincided with the upper orlower limits (limits in terms of avoidance of interactionwith counterpart parts). Consequently, as can beunderstood from Fig. 8(a), there was no apparentdifference between the L27 best solution and theoptimal values obtained from the response curvedsurface. We cannot deny the possibility of obtainingbetter optimal values, but we believe that we havesuccessfully determined values very near the idealoptimal 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, wedetermined the shape of ourfinal design.

6. Final Design

In the development work, wealso applied the analysistechnique described above tothe outer ring. In addition toshape optimization, weattempted to achieve lighteningin the finalized bearingspecification through alterationof bearing internal design anddevelopment of new materialsand grease. The newlydeveloped hub bearing shape,which is illustrated in Fig. 9, has

-18-


Fig. 10 Amount of each component Lightening

(b) Stress value

Stress point 1

L27 best solution

Stress point 2 Stress point 3

(a) Volume

Optimization L27 best solutionInitial value Optimization

Fig. 8 Result of shape optimization

Fig. 9 Shape of final design

－100g

Outer ring

Hub ring

－240g

achieved the target mass of 1.0 kg and cleared thetarget values for strength, durability and rigidity. Inaddition, the hub ring and outer ring have alsoachieved lightening targets as illustrated in Fig. 10.

7. Conclusion

This report has introduced the shape optimizationanalysis executed for lightening hub bearings forreduced weight cars. With this technique, we havesucceeded in achieving a target mass of 1.0 kg,attaining dramatic lightening that was previouslyconsidered impossible. Note, however, that this newlydeveloped hub bearing is not suitable for drum brakestructures in which the hub bearing also functions as abrake seal.* For such applications, another hubbearing type developed through use of our shapeoptimization 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 settargets for hub bearing rigidity. Therefore, we will aimfor multi-faceted optimization in which we attempt topromote further lightening while maintaining sufficientbearing rigidity so that we can establish hub bearingtechnology to cope with the needs of various carmanufacturers.

*As shown in Fig. 9, the newly developed hub bearinghas four pawls that radially guide the membersinstalled to the hub ring. If the hub bearing is appliedto a drum brake, the pawls need to be replaced with acircumferential rim to provide sealing function.

-19-

Haruo NAGATANI

Automotive Sales HeadquartersAutomotive Engineering Dept.

Photos of authors

Tsuyoshi NIWA


References1) J. Sakamoto, "Hub unit bearing market and

technology trends," Monthly Tribology, October 20042) K. Kajihara, "Improved hub unit analysis simulation

techniques," Koyo Engineering Journal No. 167(2005)


-20-


[ Technical Paper ]

Dynamic Analysis of a High-Load CapacityTapered Roller Bearing

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 rollingelements and the cage itself as well as the resultantforces occurring between the bearing components isnecessary. In short, dynamic analysis is needed.

NTN had previously developed a dynamic analysistool for cylindrical roller bearings limited to two-dimensional freedom by using the general-purposemechanism analysis software ADAMS(R).1) 2) 3) Byenhancing this dynamic analysis technique, we haverecently developed a three-dimensional dynamicanalysis tool for tapered roller bearings.4)

NTN has developed a high-load capacity taperedroller 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 inthe cage behavior and the interaction between therollers and the cage by comparing this new type andstandard tapered roller bearings using our newlydeveloped 3D dynamic analysis tool. As a result, welearned that the new type does not greatly affect thecage behavior or the interaction between the rollersand the cage.

This report summarizes the 3D dynamic analysistool and describes the results of our analysis.

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

It is necessary to predict forces acting on a cage when designingrolling element bearings. It requires a dynamic simulation that canevaluate interaction forces between the bearing componentsincluding the cage as well as real-time behaviors of thesecomponents.

NTN had already developed a 2-dimensional dynamic analysiscode for cylindrical roller bearings using a commercial, versatiledynamic analysis software. At this time, NTN has developed a 3-

dimensional 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 interactionsbetween the cage and rollers, and the experimental verification of these interactions is generally beyondaccurate measurement. Accordingly, the developed code is implemented to investigate the difference betweenthe conventional tapered roller bearings and the newly proposed one in terms of the interaction forces betweenthe 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 effectivefor use.

Kazuyoshi HARADA*Tomoya SAKAGUCHI*

Dynamic Analysis of a High-Load Capacity Tapered Roller Bearing

2. Symbols

b : Half the Hertz contact length, mD : Cage outside diameterDiso : Deborah Number =η0eαP

‾

u＿

/（Gb）G : Dimensionless material parameter =α0E'

E' : Equivalent Young's modulus, PaFa : Axial load, NFEHLr : EHL rolling viscous drag, NFpx : Rolling direction component in EHL oil film

pressure, NFr : Radial load, Nfc : Cage running frequency, HzfT : Friction force at contact area, Nhc : Oil film thickness at center, mk : =1.03αr

2/π

k' : Thermal conductivity of lubricant, W/(mk)Lt : Thermal load coefficient =η0βu

＿2/ k'

l : Width of roller slice, mNsl : Number of roller slices P―

: Mean surface pressure on Hertzian contact, PaPmax : Maximum surface pressure on Hertzian

contact, Paq : Contact force on sliced piece, NRe : Equivalent radius of curvature, ms : Slip ratioS―

: Mean dimensionless shear stress on thewhole contact area

U : Dimensionless representative velocity=η0u

＿

/（E'Re）u＿

: Mean surface velocity, m/sus : Sliding velocity on contact areaW : Dimensionless load parameterXc : Dimensionless length on EHL contact area

=（Diso/Σiso）sinh-1Σiso

xc* : Dimensionless x-direction displacement atcage mass center = x/δPr

yc* : Dimensionless y-direction displacement atcage mass center = y/δPr

zc* : Dimensionless z-direction displacement atcage mass center = z/δPa

α : Pressure coefficient of viscosity, 1/Paα0 : Pressure viscosity index of lubricant under

normal pressure, 1/Paαr : Ratio of curvature radius vertically

intersecting the rolling direction to curvatureradius in rolling direction

αp : Pocket angleβ : Temperature increase dependent

coefficient of viscosity, 1/Kδ : Geometric interference amount, mη0 : Viscosity under normal temperature and

pressure, Pa･sψ : [1+2/(3αr)]

-21-

Λ : Film thickness parameter = hc /σe

Λbd : Upper limit film thickness parameter underboundary lubrication

Λhd : Lower limit film thickness parameter under hydrodynamic lubrication

μbd : Friction coefficient under boundary lubricationμhd : Traction coefficient with oil filmμr : Friction coefficient on contact areaΣiso : Dimensionless sheer velocity with

isothermal lubricantσe : Equivalent roughness of contact areas

between two objects, mδPa : Cage axial clearance δPr : Cage pocket radial clearance τ0 : Lubricant characteristic stress, PaφT : Temperature compensation coefficient of oil film

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

3. Analysis method

The conditions assumed for our analysis aresummarized below.¡The rollers and cage are provided with six degrees

of freedom.¡The inner ring is subjected unconditionally to

translational displacement equivalent to rotation atspecific velocities and preset loads (zero degree offreedom).

¡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 intoaccount.

¡Interaction force distribution on the roller rollingcontact surface is evaluated with the slice method.

¡For the friction force between the rollers and theraceway, 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 (speed-dependent term) is not considered.

¡All the interaction force between the roller large endface and the inner ring large rib face is assumed to


-22-

be applied to the maximum neighboring point. Thefriction coefficient is handled in the same manner asthat of the raceway surface. However, because ofthe sliding contact, the EHL rolling viscousresistance 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, allthe contact force and friction force of the maximuminteraction point is applied.

3.1 Dynamic interaction forces between the roller contact surface and the raceway surface

The dynamic interaction force model for theinteraction between the roller contact surface and theraceway surface is based on the following assumptions.1) Only when an elastic deformation amount

(geometric interaction amount δ) occurs againstthe relative position of the rollers and the bearingring that is governed by the time t, a contact-induced normal force, friction force and rollingviscous resistance occur between the roller rollingcontact surface and the raceway surface.

2) The slice method is applied to the roller rollingcontact surface, thereby the distribution of theinteraction force is considered. If the interactionvector δ on each slice is given, then Palmgren'ssimple formula is employed and the expression (1)is used to determine the contact force vector q foreach slice.

3) The friction coefficient is calculated for eachlubrication mode in the manner shown byexpression (2), based on the film thicknessparameter.3) The friction force vector fT acting oneach roller slice is determined with expression (3)using the roller-dependent sliding velocity vector us.

Contact andFriction forces





Contact and Tangential (Traction or Friction) forcesEHL Rolling Resistance




Contact and Tangential (Traction or Friction) forces

Fig. 1 Considered interaction forces in this dynamic analysis

q=0.356E'Nsl-1/9 l 8/9δ10/9

δ δ ………………（1）

μr=

………………………………………………（2）

…………………………………（3）

μbd－μhd

μbd if Λ＜Λbd

if Λbd≦Λ＜Λhd

if Λhd≦Λμhd

（Λbd－Λhd）6（Λ－Λhd）6＋μhd

ft=μrq・ us

us

hc=

hc, PE=2.922 W-0.166 U0.692 G0.47 Re

hc, IR=4.9 UW-1 Re

…………（4）

…………（5）

………………………（6）

if hc, PE＞hc, IR

otherwise

φT・hc, PE

φT・hc, IR

For this purpose, Λbd=0.01 and Λhd=1.5 wereused. The above-mentioned film thicknessparameter 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), asnecessary) as described in expression (4). For thePE mode, Pan's formula (5)6) was used while for theIR 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 IRmode is necessary when the rollers are in contactwith the outer ring in the non-load region.Additionally, through comparison of the oil filmthicknesses of the two fluid lubrication modes, therelevant mode was determined, and this mode wasused to judge whether the EHL rolling viscousresistance, described later, was necessary.

-23-

For the friction coefficient μbd under the boundarylubrication condition (Λ<Λbd) shown by theexpression (2), the function9) of Kragelskii's solidcontact 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 setto 0 when sliding was also 0.

The friction coefficientμhd for the fluid lubricationcondition (Λ<Λhd) was calculated with Muraki'ssimple theoretical formula10) given below. To reducethe amount of numerical calculations, thetemperature was assumed to be constant. Inaddition, it was assumed that this traction modelcould also be effective even in the IR mode.

Here, the dimensionless length of the elastic areais represented as Xc = ( Diso/Σiso) sinh-1Σiso.

Slip ratio

Fric

tion

coef

ficie

nt

0.15

0.12

0.09

0.06

0.03

0 0.01 0.02 0.03 0.04 0.050

Oil film parameter　Λ

Fric

tion

coef

ficie

nt　μ

r

0.15

0.12

0.09

0.06

0.03

0.001 0.01 0.1 1 100

μhd =0.06

μhd =0.001

Fig. 2 Friction coefficient under boundary lubrication

Fig. 3 Relationship between friction coefficient and oilfilm parameter

The mixed lubrication region (μbd±Λ＜μhd) wasdetermined by smoothly interpolating the frictioncoefficients μhd andμbd of the above-mentionedfluid lubrication and boundary lubrication,respectively, as shown by expression (2). Fig. 3shows 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 ofEHL rolling viscous resistance was assumed to beopposite 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 directionon the EHL oil film.5) Note, however, that the sign forthe rollers and inner ring must be a plus sign, andthat for the outer ring must be a minus sign. Beingcompensated for by the component force ofpressure in the normal direction, this force does notaffect the moment.

As such, the contact force, friction force and EHLrolling viscous resistance acting on each roller slicedynamic contact surface were calculated. Inaddition, the moment on each slice was calculated.The total of forces and moments working on all theslices act on the roller.

μbd =（－0.1＋22.28s）exp（－181.46s）＋0.1…（9）

Xc≧2：S=Σiso/Diso ………………………（10）

Xc＜2：S=sinh-1Σiso｛1－（Diso/4Σiso）sinh-1Σiso｝

μhd=τ0 S/P

………………………（11）

………………………………（12）

FEHLr=

Fpxb,r=±

………………………………………………（13）

…………………………（14）

otherwise

if hc, PE＞hc, IR－φT α0

Rb,r

2Re

29.2Rel（GU）0.648 W0.246

0

・ uu

uu

FEHLr

………（8）

…（7）

for Roller / Inner race

for Roller / Outer race

φT =

Provided that,

Rb

1

Rb

1

Ri

1

Ro

1

＋

－

1＋0.213（1＋2.23s0.83）LT0.64

1－13.2（Pmax /E'）LT0.42

Re-1=


Because the roughness of the pocket surface is great,the friction force when the roller rolling contact surfaceis in contact with the cage pocket surface wascalculated assuming that the lubrication mode wasboundary lubrication. In addition, because the width ofthe bar of the pocket in contact with the roller rollingcontact 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 endface and the cage pocket surface

The interaction forces between the roller large endface 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 therepresentative points where contact could occur to aseries of calculations, and took the sum of the contactand friction forces at each point as the interactionforce at that point. Because the roughness of thepocket surface is great, the friction force when theroller end face was in contact with the cage pocketsurface was calculated assuming that the lubricationmode was boundary lubrication.

4. Analysis model

The analysis model and coordinate systems usedare shown in Fig. 5. Every coordinate system usedwas a right-hand coordinate system. Two types ofbearings "standard specification and high-loadcapacity" were subjected to dynamic analysis todetermine the differences in the cage behavior andinteraction force between them. The high-loadcapacity type has inner and outer rings and rollerswhose diameters were the same as those on thestandard specification type, but its newly developedcage has an outside diameter greater than that of thestandard specification bearing. The bearing data andoperating conditions for these bearing types aresummarized 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 ringwas displaced in the +y direction and the topmostroller received the maximum load.

3.2 Interaction forces between the roller largeend face and the inner ring large rib face

Generally, the contact area between the roller largeend 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 round-

shaped and the inner ring large rib face is cone-shaped. As a result, elliptical contact occurs.Because the length of elliptical contact is smallerthan the contact length between the roller and theraceway surface, it can be assumed that all theinteraction forces act on the geometrical maximuminteraction point.

2) From the geometrical interaction amount δ, thecontact force can be calculated with Hertz's pointcontact formula. The friction coefficient for the rollerlarge end face and inner ring large rib face can becalculated in a manner identical to that for theraceway 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.

3.3 Interaction forces between the roller smallend face and the inner ring small rib face

Generally, contact between the roller small end faceand 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 interactionforce is applied to the maximum contact point, wecalculated the interaction force using Hertz's pointcontact formula. Because the mode of contact in thisarea was considered to be edge contact, we used afriction coefficient under constant boundary lubricationfor the friction force calculation.

3.4 Interaction forces between the dynamic rollercontact surface and the cage pocket surface

The interaction force between the dynamic rollercontact 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-


Detail of pocket

Large endside

Small end side

Fig. 4 Geometrical shape of cage pocket

hc=

hc, IR=128αr 0.131tan-1

……………（15）

……（16）

…（17）

if hc, PE＞hc, IR

otherwise

φT・hc, PE

φT・hc, IR

αr ψ 2

2 2U

W＋1.683 Re

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

-25-

Fig. 5 Analyzed bearing schematic and its coordinate system

Fig. 6 Cage geometry

Table 1 Test bearing and operating conditions

5. Analysis results

In further analysis, to eliminate the influence of theinitial conditions the data acquired during the 0.5 stime-span after the start of calculation (forapproximately 40 revolutions of the inner ring) wasdeleted, and the data obtained during the next 0.2 s(for approximately 17 revolutions of the inner ring) wasutilized for evaluation.

5.1 Translational displacement on the radialplane at the cage mass center

On the cage used for the analysis, the pocket radialclearance is larger than the pocket circumferentialclearance as illustrated in Fig. 6(b), and the maximumtranslational displacement of the cage is the radialclearance on the pocket 4). The trajectories of the cagemass centers on their radial planes with the standardbearing and the newly developed bearing are plottedin Fig. 7. The red solid lines in the charts representthe radial plane positions of the cage mass centersthat were made dimensionless with the radial pocketclearances. The blue dotted lines indicate circles thattook the radial pocket clearances of the cages as aradius. The areas of trajectories of cage mass centersare smaller than the radial pocket clearances, andmatch the characteristics of cage behavior undercomplex load conditions described in previousresearch4, 13). The displacement of the newlydeveloped cage is slightly smaller than that of thestandard cage.

The pattern of interaction force between the rollerrolling contact surface and the cage is illustrated inFig. 8. Fig. 8(a) shows the cage/roller interaction forceand the interaction position in the case where thecage mass center is displaced to the position shown inFig. 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,

Bearing (inside dia.×outside dia.×width, mm)

High-loadcapacity bearing*

Standard specificationbearing

（φ40×φ76.2×17.5）

Cage typeNumber of rollers

Basic rated dynamic load, kNPocket angle αp, deg

Cage outside dia φD，mmLubricant temperature ˚C

Lubricant dynamic viscosity, mm2/s {cSt}Inner ring running speed, rpm

Load, kN

Standard cage21

46.545

61.39

Newly developed cage23

50.056

62.22

1002.52 @ 100˚C

5000Fr = 5，Fa = 2.5

*The inner ring, outer ring and rollers are identical to those used on the standard specification bearing.

φD

Pocketcircumferentialclearance

Poc

ket r

adia

lcl

eara

nce

(a) Assembled bearing

αp

(b) Cross-sectional viewof cage pocket

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

1.0

0.5

0.0

-0.5

-1.0-1.0 -0.5 0.0 0.5 1.0 -1.0 -0.5 0.0 0.5 1.0

1.0

0.5

0.0

-0.5

-1.0

Displacement xc＊

Radial Pocket Clearance

Displacement xc＊

Dis

plac

emen

t y c＊

Dis

plac

emen

t y c＊

(a) Standard cage (b) Newly developed cage


under 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 withinthe 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 forcebetween the dynamic roller contact surface and thecage is given in Fig. 9 for both the standard cage and

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

-26-


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

100.0 100.0

0.0 0.00.5 0.50.6 0.60.7 0.7

Time, s

Cag

e/R

olle

r F

orce

, N

Cag

e/R

olle

r F

orce

, N

Time, s


50 0.3

0.25

0.2

0.15

0.1

0.05

0

45

40

35

30

25

20

15

10

5

0Currentbearing

Developedbearing

Maximum forceMean force

Max

imum

forc

e, N

Mea

n fo

rce,

N

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

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

-1.0 -0.5 0.0 0.5 1.0

1.0

0.5

0.0

-0.5

-1.0Position of cage mass center

Displacement xc＊

Dis

plac

emen

t y c＊

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

x

y

has decreased due to the increased number of rollers.The greater number of rollers has caused the tractionforce from the raceway surface that drives the rollersto decrease and resulted in decreasing the meaninteraction force between the dynamic roller contactsurface and the cage.

The driving force from the cage stems from theinteraction force on the rollers, and the time-integration value of this interaction force governs thebehavior 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 was2%. 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 interactionforce with the newly developed cage did not exceedthat of the standard cage by more than 2%. Thus, wecan judge that the maximum interaction force isroughly the same for both cage types.

5.2 Axial displacement of the cage mass centerNext, the time-dependent trend of the axial

displacement of the mass centers of both cages isillustrated in Fig. 11. As shown in Fig. 12, the verticalaxis took the cage mass center as the zero point whenthe 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 axialclearance 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 thecage. In Fig. 13, cage/roller interaction force patternsin the cage pocket are illustrated. The red arrow in thediagram shows the interaction force from the cageacting on the roller. As a result, a reaction force inresponse to the illustrated interaction force acts on thecage. Due to the geometrical shape of the cagepocket, 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 ofthe above-mentioned interaction force. On the otherhand, when the roller large end face interferes with thecage pocket (Fig. 13(b)), the cage shifts in the +zdirection.

-27-

Fig. 11 Axial behavior of cage mass center

1.0

0.8

0.6

0.4

0.2

0

1.0

0.8

0.6

0.4

0.2

00.5 0.6 0.60.55 0.50 0.550.7 0.70.65 0.65

Time, s

Dis

plac

emen

t, z

c＊

Dis

plac

emen

t, z

c＊

Time, s


zc

Contact areas

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


The roller large end face/cage pocket interactionforce is summarized in Fig. 14. This chart wasobtained 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 frequencyfc 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 newlydeveloped cage. In terms of axial displacement, thedifference between the standard cage and the newlydeveloped cage in Fig. 11 results from the differencebetween interaction patterns given in Fig. 14. Underour analysis conditions, no collision between the rollersmall end face and the cage pocket occurred.

As described above, the behavior of the newlydeveloped cage is the same as that of the standardcage, while the roller/cage interaction force with thenewly developed cage is smaller than that with thestandard cage. The effect of the cage outsidediameter 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 standardtapered roller bearing and a high-load capacitytapered roller bearing that has an increased number ofrollers. As a result, we have found that the behavior ofthe cage with the high-load capacity bearing is verysimilar to that of the standard bearing and that theroller/cage interaction force with the high-load capacitybearing 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-loadcapacity bearing onto the roller/cage interaction forceis small within the scope of our analysis.

It is difficult to experimentally compare the levels ofroller/cage interaction force. Therefore, our analysistechnique, as an alternative to an experiment-basedtechnique, is a useful means for verifying the functionsof high-load capacity bearings.

-28-


(a) Contact between roller rolling surface and cage pocket (b) Contact between roller large end and cage pocket

Cagerevolutiondirection

Cagerevolutiondirection

Z

XZ

X

Interaction force fromcage acting to roller

Interaction forcefrom cage actingto roller

Lar

ge

end

fac

e

Sm

all e

nd

fac

e

Sm

all e

nd

fac

e

Lar

ge

end

fac

e

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

Fig. 14 Cage/roller large end interaction forces

10.0

7.5

5.0

2.5

00.5 0.60.55 0.70.65

Time, s

0.5 0.60.55 0.70.65

Time, s

Cag

e/R

olle

r-en

d fo

rce,

N

10.0

7.5

5.0

2.5

0Cag

e/R

olle

r-en

d fo

rce,

N


References1）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 rollerbearings, Japanese Society of Tribologists, TribologyConference Preprints (Sendai 2002-10) 415.

3) Tomoya Sakaguchi, Kaoru Ueno: Analysis of cagebehavior on cylindrical roller bearings, NTNTechnical 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 TaperedRoller Bearings, Trans. ASME, J. Trib., 113, 7（1991） 590-597.

6）Pan, P., Hamrock, B.J.: Simple Formulae forPerformance Parameters Used inElastohydrodynamically 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.

-29-

Kazuyoshi HARADA

New Product DevelopmentR&D Center

Mechatronics Research Dept.

Photos of authors

Tomoya SAKAGUCHI




8）Gupta, P. K. et al., : Visco-Elastic Effects in Mil-L-7808 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 intotraction 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 isothermalelastohydrodynamic lubrication of concentratedcontacts I. Direction of lubricant entrainmentcoincident with the major axis of the Hertzian contactellipse, Proc. Roy. Soc., London, A397（1985）245-269.

13) Kazuyoshi Harada, Tomoya Sakaguchi: Analysis ofcage behavior on tapered roller bearings (1st report,behavior measurement), Japanese Society ofTribologists, Tribology Conference Preprints (Tottori2004-11) 501.

-30-


[ New Product ]

High Capacity Tapered Roller Bearings- Super Low Torque・・High Rigidity Tapered Roller Bearings -

1. Introduction

In order to reduce automobile fuel consumption,efforts are in progress to introduce low-viscosity oilsand compact, lightweight designs for automotivetransmissions and differential gears. The sizes ofbearings used in transmissions and differential gearshave been getting much smaller, and, as a result, theneed to ensure bearing life and rigidity has beenincreasing. To address this challenge, we havedeveloped a unique high-load capacity, compacttapered 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 samenumber as a full complement roller bearing, toincrease the load carrying capacity. As a result, thecontact pressure is decreased, higher bearing rigidityis achieved and bearing life is extended even undersevere 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.

Tapered roller bearing have greater capacity for carrying not only pure radial or axial loads but alsocombined loads, and feature greater bearing rigidity. Therefore, they are found in numerousapplications in various industries such as the automotive industry.

Recent advancements in transmissions for low fuel consumption have resulted in lower oilviscosities and reduction in transmission size. Therefore, it is necessary to reduce the size of thebearings, 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 loadrating can be increased by up to 10%, and the static load rating can be increased by up to 15%. Asa 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 combininghigh capacity tapered roller bearings and FA tapered roller bearing features.

Takashi TSUJIMOTO*Jiro MOCHIZUKI*

High Capacity Tapered Roller Bearings

2. Structure of high-load capacitytapered roller bearing

Employing a smaller clearance between the cageand the outer ring to increase the cage PCD allowsthe cage bar width to remain the same as that of a

-31-

Clearance betweenouter ring and cage

Clearance betweenouter ring and cage(smaller)Cage bar width

Cage bar width(same as conventional bearings)

Clearancebetween rollers

Clearancebetween rollers(smaller)

Standard bearing High-load capacity bearing

Fig. 1 Structure of high capacity tapered roller bearing

Standard bearing High-load capacity bearing

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%)

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

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

3. Features of high-load capacitytapered roller bearing

The increased number of rollers gives our high-loadcapacity tapered roller bearing the following improvedfunctions:

1) Increased rated load¡Up to 10% increase in dynamic rated load

(maximum improvement of 37% in terms ofcalculated life)

¡Up to 15% increase in static rated load (maximumimprovement 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

lubricating conditionsThe greater number of rollers contributes to

reducing the maximum contact surface pressure,increasing the oil film thickness, and stress reliefin metal-to-metal contact situations. Thus,surface-initiated flaking, which results from metal-to-metal contact in lubricating conditions when anoil film is not readily formed, is inhibited resultingin 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. Asa result, the bearing life under contaminatedconditions is also extended.

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


-32-

4. Performance of high-load capacitytapered roller bearing

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

Fig. 3 illustrates the results of a bearing life testunder Λ*0.2 clean oil lubricating conditions in whichoil 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 lifeapproximately 15 times longer than standard taperedroller bearings.

9590807060504030

20

10

5

2 2 21 1 13 35 57 7×100 ×101 ×102

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y,

%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: Turbine Oil VG56Contaminants: steel beads + gas-atomization 0.5 g/L Contaminant grain size 50μm or smaller, 90wt% 100-180μm, 10wt%Lubrication method: oil bathOil film parameter: 　= 2.0Calculated life: standard bearing L10= 92.2 h high-load capacity bearing L10= 120.4 h

Standard bearingHigh-load capacitybearing

1.9713.275

slope e

22.549.0

L (50)

8.727.6

L (10)

Λ

Fig. 4 Life test results under contaminated lubrication

9590807060504030

20

10

5

2 2 21 1 13 3 3 35 5 5 57 7 7 7×100 ×101 ×102

1×103

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y,

%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 VG1.5Lubrication method: oil bathOil film parameter:　 = 0.2Calculated life: standard bearing L10= 92.2 h High-load capacity bearing L10= 120.4 h

Standard bearing

1.5464.231

slope e

22.4152.9

L (50)

6.698.0

L (10)

High-load capacitybearing

Λ

Fig. 3 Life test results under clean lubrication

Table 1 Test bearing

High-load capacity bearingStandard bearing

Bearing size

Rated load

Number of rollers

φ45×φ81×16 mm

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

Dynamic rated load Cr =45.5 kN Static rated load Cor=58.0 kN

24 rollers 27 rollers

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

2) Results of bearing life test for use withcontaminated lubricant

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 lifeapproximately 3 times longer than standard taperedroller bearings under the same conditions.3) Results of bearing life test in use with low-

viscosity lubricantGenerally, 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 onthe market are VG20 level oils. Because transmissionoils with much lower viscosity will be increasinglyused, we executed a bearing life test using viscositygrade VG10 oil under both clean oil lubrication andcontaminated oil lubrication conditions.

-33-

9590807060504030

20

10

5

2 21 13 35 577 7×101 ×102

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y,

%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 bathOil film parameter: 　= 0.8Calculated life: standard bearing L10= 92.2 h High-load capacity bearing L10= 120.4 h

Standard bearingData for samples whose tests were terminated before they failed

High-load capacitybearing

1.7623.328

slope e

146.0188.1

L (50)

50.1106.8

L (10)

Λ

Fig. 5 Life test results under clean lubrication

9590807060504030

20

10

5

1 1 15 5 5 5×100 ×101 ×102

1×103

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y,

%

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)Contaminants: steel beads + gas-atomization 0.3 g/L Contaminant grain size 50μm or smaller, 90wt% 100-180μm, 10wt%Lubrication method: oil bathOil film parameter: 　= 0.8Calculated life: standard bearing L10= 92.2 h high-load capacity bearing L10= 120.4 h

Standard bearingHigh-load capacitybearing

1.6851.158

slope e

49.8237.2

L (50)

16.346.6

L (10)Λ

Fig. 6 Life test results under contaminated lubrication

The results of the life test under the Λ=0.8 clean oillubrication 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 formationconditions and boasts a bearing life approximatelytwice as long as standard tapered roller bearings.

The results of the life test under the Λ=0.8contaminated oil lubrication condition are shown inFig. 6.

We performed a market study of the contaminantsin lubricating oils and found the size of contaminantswas 50μm or smaller and that the amount ofcontaminants was less than 0.3 g/L even on MT cars.Therefore, we employed severe contaminationconditions 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 duringuse with contaminated lubricant and boasts a bearinglife approximately 3 times the length of the standardtapered roller bearing.

5. Application to extremely lowtorque/high rigidity design

Our high-load capacity tapered roller bearing can becombined with a long-life bearing, such as the NTN FAtapered roller bearing 2) that features a special heattreatment process (FA), to obtain finer crystal grainsand optimized bearing interior design technologies.

The FA tapered roller bearing offers longer bearinglife, greater seizure resistance and improved dentresistance. If it is designed for a bearing life equivalentto standard tapered roller bearings, its size can bemore compact than that of standard bearings,resulting in lower torque. However, loss in bearingrigidity was unavoidable when we attempted toachieve a significantly lower torque. By combining theFA tapered roller bearing with our newly developedhigh-load capacity tapered roller bearing, it is possibleto avoid loss in bearing rigidity and achieve a lowertorque.


Based on this finding, we have summarized thetorque factors and contribution rates of a tapered rollerbearing used on a pinion shaft support in a near axledifferential 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 causedby lubricant viscosity," "rolling viscous torque on rollingcontact surface" and "shear torque on lubricantbetween cage and rolling elements."

Furthermore, it is possible to decrease the torque to50% by greatly reducing the lubricant stirring drag inthe bearing through optimization of the cage shape.Applications of 50% low torque design that alsoensure bearing life and rigidity are described below.

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 forlow-torque tapered roller bearings is growing.

To address this problem, we have attempted toachieve lower torque for a bearing used for pinion shaftsupport in near axle differentials with model #30306D.

5.2 Torque factors on tapered roller bearingsThe torque factors on tapered roller bearing are

shown in Fig. 7.To decrease the torque occurring on tapered roller

bearings, factors 1 through 5 need to be reduced.

-34-


Fig. 7 Torque factor of tapered roller bearing

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

Among the torque factors working on a taperedroller 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 inFig. 9, to determine the magnitude of stirring torqueon tapered roller bearings used for pinion shaftsupport 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.

1 Rolling viscous torque on rolling contact surface

5 Stirring torque caused by viscosity of lubricant

3 Torque resulting from elastic deformation of rolling elements

4 Shear torque on lubricant between cage and rolling elements

2 Sliding torque on rib surface

Fig. 9 Torque measurement method

Low oil level

High oil level

Fig. 10 Factors contributing to friction torque of tapered rollerbearing used for pinion shaft support in near axle differential

30%57%

11%

2%0%

Rolling viscoutorque on rollingcontact surface

Sliding torque on rib surfaceTorque resulting from elasticdeformation of rolling elements

Shear torque on lubricantbetween cage and rollingelements

Stirring torquecaused bylubricant viscosity

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

Rot

atio

nal t

orqu

e，N･c

m

200

160

120

80

40

01000 2000 3000 4000 50000

Bearing speed, r/min

High oil levelLow oil level

Load conditions: axial load 6000 NLubricant: hypoid gear oilLubrication method: oil bath (natural circulation)

30% torquereduction

-35-

5.4 Torque reduction techniquesReduction of the major contributing torque types is

necessary to decrease the torque on tapered rollerbearings used on pinion shaft support in near axledifferentials. These include "rolling viscous torque onrolling contact surface", "shear torque on lubricantbetween cage and rolling elements" and "stirringtorque caused by lubricant viscosity".

1 Decrease in rolling viscous torque on rollingcontact surface

Optimization of the bearing interior design and thecrowning of rollers is effective for decreasing therolling viscous torque on a rolling contact surface. Ourtorque reduction techniques that can ensure bearingrigidity are described below.

1) Alteration of bearing internal designTo achieve a lower torque while maintaining

bearing rigidity, we have checked the bearinginternal design's contribution to torque and bearingrigidity. The results are summarized in Fig. 11. Toachieve a high-rigidity low-torque design, smallerroller pitches and greater contact angles areadvantageous.

2) Alteration of crowning shapeFig. 12 summarizes the torque reduction

(calculated values and actual measurements) thatresult from the 17% reduction in the effectivecontact length between the rollers and the racewaysurface through alteration of the crowning shape.The torque can be decreased by altering thecrowning shape to reduce the effective contactlength. The actual effect nearly matches ourcalculated values.

However, as summarized in Fig. 13, a lowertorque attained by alteration of the crowning shapecan decrease the bearing rigidity. Therefore,bearing rigidity must be considered to determine an

Fig. 11 Effect of internal design factors in torque and rigidity

Degree of effect1 1.5 20.50-0.5-1-1.5-2

Larger rollerdiameter

Larger rollerpitch

Increased number ofrollers

Largercontact angle

Longer roller

Torque reduction effectHigh rigidity effect

Fig. 12 Torque reduce percentage by changing crowning

Torq

ue r

educ

tion

%

30

25

20

15

10

5

00 1000 2000 3000 4000 5000

Bearing speed r/min

Calculated valuesActual measurements

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

Torq

ue r

atio

, rig

idity

rat

io

1.3

1.2

1.1

1.0

0.9

0 0.5

1.51 2

0.8

0.7

Inner ring crowning radius ratio

Crowning shape examples

Torque ratioRigidity ratio

optimal crowning shape.

2 Reduction of shear torque on lubricantbetween cage and rolling elements

As summarized in Table 2, compared to a standardcage, 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 promoteoil flow toward the outer ring. This arrangement canlower the lubricant stirring torque compared to caseswhere lubricant oil tends to remain stagnant in thebearing inner ring.


3 Decrease in stirring torque caused byviscosity of lubricant

1) Effect by modified bearing internal designA low torque design with a more compact bearing

results in reduced roller rolling area volume asshown in Fig. 14, thereby reducing the lubricantstirring torque inside the bearing. An example ofreduced roller rolling area volume is describedbelow.

Fig. 15 shows a comparison between thestandard bearing and a new bearing with a modifiedinternal design. As summarized in Table 3, therolling area volume in the bearing with a modifiedinternal design is 31% smaller than the standardbearing.

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 lowviscosity of the lubricant, the reduction effect in thecalculated values is lower than the actualmeasurements. However, when the reduction in thelubricant stirring torque due to the roller rolling areareduction is considered, as demonstrated byexpression (1), the calculated reduction in bearingtorque coincides well with the actual measurement.

-36-


Torquereductionpercentage

―

0%

10%

10%

12%Oil flow

10%

Standard cage

Smallerdiameterside

Largerdiameterside

Shape A Shape B

Calculatedvalues

Actualmeasurements

Table. 2 Type of cage and torque reduction ratio

Fig. 15 Comparison of bearing formFig. 16 Relationship between torque and torque reduce

percentage

Table. 3 Comparison of bearing internal designRoller rolling area

volume ratioRoller pitch dia.Roller length Avg. roller dia.

Standard bearing

Bearing withmodified internaldesign

13

11

9.22

8.04

51.54

48.46

1.00

0.69

Standard bearing New bearing with modifiedinternal design

20.75 20.75

14 14

19 19

φ30

φ72

φ30

φ72

Fig. 14 Roller rolling area volume

Bearing torque＝30%×（100%－31%）＋70%

　　×（100%－13%）＝82%（18% reduction）…Expression (1)

Lubricant stirringtorque rate

Roller rolling areavolume reduction rate

Rolling viscositytorque reduction rate

Rolling viscositytorque rate

Calculated value coincides well with actual measurement.

Torq

ue r

educ

tion

%

30

25

20

15

10

5

00 1000 2000 3000 4000 5000

Bearing speed r/min

Approx. 5%18%

13%

Calculated valueActual measurement

2) Alteration to cage inside diameterTo reduce the lubricant stirring drag by decreasing

the amount of lubricant flowing into the bearing, weverified 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 lubricantstirring torque accounts for approximately 30% ofthe whole bearing torque. In other words, thelubricant stirring torque was reduced by 83%.

-37-

Table 4 Bearing internal design

Low-torque designStandard bearing

Rated load

Roller PCD (mm)

Number of rollers

Contact angle

Roller length (mm)

Mean roller dia. (mm)

Mass (kg)

Dynamic rated load Cr =49.0 kNStatic rated load Cor=52.5 kN

Dynamic rated load Cr =33.0 kNStatic rated load Cor=35.5 kN

φ51.54 φ44.44

15

28°48′39″

17

30°30′

13 10.1

φ9.22 φ7.25

0.393 0.223

Fig. 17 Relationship between speed and torque reducepercentage of bearing with small diameter cage

Fig. 18 Comparison of bearing size

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

40

30

20

10

00 1000 2000 3000 4000 5000

Standard cage Small diameter cage

Torq

ue r

educ

tion

%


1) Torque reduction effect by factor(calculated value)Fig. 19 summarizes the running torque

reduction effect (expected value) for eachtorque reduction factor. In the practicalspeed range of 2000 to 3000 r/min, it ispossible to reduce running torque by 50%.

Standard bearing Low-torque bearing

20.75 10.4

16

14

19 16φ30

φ72

φ30

φ60

More compactbearing designis possible.

Run

ning

torq

ue r

educ

tion

%

60

40

50

20

30

00 1000 2000 3000 4000 5000

10


1

2

3

4

5

1 Effect of modified internal design2 1+crowning effect 3 2+effect of reduction in shear torque on lubricant between cage and rolling elements4 3+effect of reduction in lubricant stirring toque by modified internal design5 4+effect of reduction in lubricant stirring torque by limiting the amount of lubricant


5.5 Ultra-low torque/high-rigidity tapered rollerbearing design example

By combining the FA bearing, the high-load capacitydesign, and the torque reduction techniques, it ispossible to reduce bearing torque by 50% whileensuring bearing life and rigidity (Fig. 18, Table 4).

2) Torque reduction effect (actual measurement)Fig. 20 summarizes the running torque

measurement results for the standard and low-torquebearings.

In the practical speed range of 2000 to 3000 r/min,running torque was reduced by 50%.

4) Axial rigidity measurement resultsFig. 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.

-38-


Fig. 20 Relationship between speed and torque of currentbearing and low torque bearing

Fig. 22 Deformation of axial direction

Fig. 23 Analysis of torque reduction

Fig. 21 Relationship between speed and torque of low torquebearing

3) Reduction effect on running torque(calculated and actual values)As shown in Fig. 21, the actual torque reductions

measured at various speeds match the correspondingcalculated values well.

Run

ning

torq

ue r

educ

tion

%

300

200

250

100

150

10000 2000 3000 4000 5000

50

0

Bearing speed r/min

50% reduction

Low-torque bearing

Standard bearing

Axi

al e

last

icity

μm

60

40

50

20

30

50000 10000 15000 20000 25000 30000

10

0

Axial load N

Low-torque bearing

Standard bearing

Run

ning

torq

ue r

educ

tion

%

60

40

50

20

30

10000 2000 3000 4000 5000

10

0

Bearing speed r/min

Expected valueActual measurement

5) Torque factor analysis for ultra-lowtorque/high-rigidity tapered roller bearing

Fig. 23 summarizes the torque values obtainedfrom our verification efforts for the ultra-low torque,high-rigidity tapered roller bearing. The NTN ultra-lowtorque, high-rigidity tapered roller bearing achieved50% reduction in bearing torque while ensuring a levelof life and rigidity equivalent to that of the standardbearing.

30%

57%80%

11%

11%5%

0%

2%

4%

0%

Standard bearing Low-torque design

Stirring torque from lubricant viscosity

Shear torque on lubricant between cage and rolling elements

Torque from elastic deformation of rolling elements

Sliding torque on rib surface (0%)

Rolling viscosity torque on rolling contact surface

6. Conclusion

As explained in this thesis, we believe that our high-load capacity tapered roller bearing, which boastslonger life and higher rigidity, offers a solution toproblems arising from challenges related to improvingautomobile fuel efficiency. We will market long-life,high-rigidity, low-torque tapered roller bearingproducts that combine FA and high-load bearingtechnologies to contribute to lower automobile fuelconsumption.

References1) Tsutomu Ohki, Kikuo Maeda, Hirokazu Nakashima:

NTN Technical Review "Longer life with bearing steelby finer crystal size," No. 71 (2003), p2.

2) NTN Catalog "FA Tapered Roller Bearings" (CAT. No.3802/E）.

-39-

Takashi TSUJIMOTO

Automotive Sales HeadquatersAutomotive Engineering Dept.

Photos of authors

Jiro MOCHIZUKI

Automotive Sales HeadquatersAutomotive Engineering Dept.


-40-


[ Technical Paper ]

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

1. Introduction

Thrust needle roller bearings, boasting variousadvantages such as high-load carrying capacity, highrigidity and compactness, are used in a diversity ofapplications. Their operating conditions andperformance requirements vary from application toapplication. Thrust needle roller bearings used for carair-conditioner compressors use a fluid mixtureconsisting of coolant and refrigerator oil as a lubricant.Recently, lower viscosity fluids are being used toimprove the efficiency of compressors, and, as aresult, lubrication performance has been decreased.To help solve this problem, long-life thrust needleroller bearings capable of low viscosity lubrication areincreasingly needed.

The authors have investigated the factors that affectthe lifespans of thrust needle roller bearings under lowviscosity lubrication conditions and have attempted tofind measures to ensure longer bearing life.

Symbols

b Half width of contactdr Roller diameterE' Coefficient equivalent to Young's modulus EE Elastic compression workload per time unit Fa Friction force occurring on protruded contact

areaFr Rolling viscous resistanceFt Traction force occurring on oil filmG Material parameterj Roller row No.L Contact lengthlrd Length of rolling contact per time unitn Number of roller rowsME Friction torque resulting from elastic

hysteresis lossMr Frictional torque from rolling viscous

resistanceMs Frictional torque from spinning

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

This paper investigates the failure modes of thrust needle roller bearings lubricated with lowviscosity lubricants both experimentally and theoretically, and proposes a new design of a long-lifethrust 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 likelycause 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 lifeas well as lower bearing rotational torque when compared with standard thrust needle roller bearings.

Hiroki FUJIWARA**Kenji TAMADA**

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

P LoadR Equivalent radiusU Velocity parameterv Absolute value of relative velocity between

race and rollersW Load parameterw Load per unit lengthw0 Load at minute width in contact centerx Coordinates in roller revolution direction

relative to revolution center as originZ Number of rollers per rowα0 Viscosity-pressure coefficientβE Elastic hysteresis loss coefficientδ Elasticity approach amountμ Friction coefficientΦE Heat generation ratio resulting from elastic

hysteresis lossφ Elastic compression workload per distance

unitφT Compensation coefficient for rolling viscous

resistance resulting from shear heat generationω Angular velocity of turning ring

2. Experimental investigation intofailure causes1)

2.1 Life test methodFor this life test, the authors employed the NTN

thrust needle roller bearing life test rig. This rig isschematically illustrated in Fig. 1, and the testconditions used are summarized in Table 1. The racemeasures 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 (surfaceroughness 0.084 μmRa). Thus, the lubrication usedfor the test was boundary lubrication.

2.2 Effects of race materials and heat treatmentfurnaces

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

-41-

Table 1 Life test condition

Fig. 1 NTN thrust needle roller bearing life test rig

Fig. 2 Effect of race material and heat treatment furnaceon bearing life

Bearings being tested

Driving motor

Hydraulicpressure

Test rig

Load

Bearing speed

Lubricant

Oil film parameter

Lubrication method

NTN thrust needle roller bearing life test rig

9.8 kN

5000 min-1

Spindle Oil ISO VG2 (70˚C)

0.1

Circulating lubrication

(a) Life test operating conditions

(b) Main characteristics of bearings used in life test

φ60×φ85×1 mm

φ3×7.8 mm

Race

Rollers

10

12

1416

18

20

8

64

2

0

Con

tinuo

us c

arbu

rizin

gfu

rnac

e

Con

tinuo

us c

arbu

rizin

gfu

rnac

e

Liqu

id c

arbu

rizin

gfu

rnac

e

Bat

ch fu

rnac

e

Con

tinuo

us c

arbu

rizin

gfu

rnac

e

Liqu

id c

arbu

rizin

gfu

rnac

e

Bat

ch fu

rnac

e

SPCC SCM415 SK5

L10

life

, hou

rs


-42-

2.3 Effect of race accuracyWe investigated the effect of race surface

roughness on bearing life. We assumed that the racesurface roughness was in a range of approximately0.03-0.11μmRa. The results are summarized in Fig.3. No significant relation between surface roughnessand bearing life was found under these testconditions.

Fig. 4 illustrates the results of investigation into theeffects of heat treatment on the deformation of theraces. The magnitude of heat treatment-induceddeformation was evaluated based on the flatness ofthe races. Deformation had no apparent effect onbearing life.

50

40

30

20

10

00.02 0.04 0.06 0.08 0.10 0.120

L10

life

in h

ours

Surface roughness (Ra) μm

Fig. 3 Effect of race roughness on bearing life

50

40

30

20

10

031 51 102 103

L10

life

in h

ours

Flatness μm

Fig. 4 Effect of race flatness on bearing life

50

40

30

20

10

00 7.3

L10

life

in h

ours

Grain boundary zone thickness μm

Fig. 5 Effect of race grain boundary oxidizing zone onbearing life

Photo 1 Raceway after life test

2.4 Effects of race materialsAs shown in Fig. 2, variation of the race steel

material and heat treatment furnace type made nodifference on the life of the thrust needle rollerbearing. We believe the reason that no differenceoccurred 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 zonewas in a range of 1.5 to 9.4μm. To determine theeffect of the grain boundary oxidizing zone, a race

Flaking

Area A

Area B

Enlarged viewof area A

Enlarged viewof area B

Rolling mark

whose grain boundary oxidizing zone was removedwith emery paper and buff-finished with diamondpaste was subjected to the life test, and the result wasas shown in Fig. 5.

The bearing life was not affected by the surfaceroughness of the race, which was 0.062μmRa beforefinishing and 0.027μmRa after buffing.

Furthermore, we investigated the possibleconnection between lifespan and other factors,including surface layer hardness, surface residualstress and residual austenitic amount, and found nosignificant relationships.

2.5 Failure typesThe appearance of the raceway that underwent the

test is shown in Photo 1. Most of the bearing failureswere shallow flakings that started at the inner boreedges 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 onthe pitch circle with slip occurring towards both endsof each roller. The bearing life might have beendecreased by slip-induced heat generation.

-43-

100

80

60

40

20

0

L10

life

in h

ours

Straight roller Crowned roller

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

0

-100

-200

-300

-400

-500

-600

-700

0.010 0.030.02 0.04 0.060.05

Res

idua

l str

ess,

GP

a

Distance from surface mm

Inner edge

Middle

Outer edge

Fig. 6 Compressive residual stress in a race after test

5 5.5 6 6.5 7

Middle

Middle

New roller

New roller

Inside edge

Inside edge

Outside edge

Outside edge

X-ray half-value width deg

Str

aigh

t rol

ler

Cro

wne

d ro

ller

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

0 50 100 200 250150 300

X-ray anisotropy MPa

Middle

New roller

Inside edge

Outside edge

Str

aigh

t rol

ler

Middle

New roller

Inside edge

Outside edge

Cro

wne

d ro

ller

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

These tested rollers and races did not use crowning,and edge loading was considered one factor forfailure. If the edge loads were a major failure factor,the failure should be internally started flaking.However, the failure that occurred in our test wasflaking 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.

2.6 Experimental verification of assumed failure causes

To verify our assumptions described above, weintroduced crowned rollers into the bearing beingtested so that the actual roller contact length wasshorter 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.

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 wetested. 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 greaterheat 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 anisotropymeasurements on the inside and outside edges andthe middles of the rolling marks on the races of bothtested bearing types. "X-ray anisotropy" can bedefined as a phenomenon where the crystals areoriented in a particular direction due to the plasticfluidization of the surface due to tangential force2).This can be qualitatively described by using theabsolute value of the difference between the residualstress measured on the rolling mark surface in theroller rolling direction and the residual stress


The friction coefficientμ(x) can be given by a mixedlubrication theory. The frictional force on a solid-to-solid contact area was calculated using theGreenwood-Tripp theory 4), and the traction force onthe fluid lubrication area was calculated using Muraki-Kimura's analysis method. 5)

Incidentally, it is usually difficult to correctly providevarious parameters that represent the surface profileand lubrication state of a solid-to-solid contact model.

Therefore, with a thinner oil film and a greaterproportion of solid-to-solid contact, friction coefficientsand frictional torques do not necessarily coincidequantitatively with experiment results.

3.1.2 Rolling viscous resistanceAccording to Zhou-Hoeprich 6), shear stress on

lubricant at the EHL inlet opposite the rolling directionoccurs on the surface of the contact object. Ifconverted into a surface integral, this shear stress canbe evaluated as a force that causes a momentopposing the revolution to occur. This force is knownas "viscous resistance." As a result of analysis of theEHL theory, Zhou-Hoeprich developed a recurrenceformula for rolling viscous resistance, as shown inexpression (2), which we adopted.

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

For thrust needle roller bearings, the frictionaltorque Mr resulting from rolling viscous resistance canbe expressed as:

n

Mr＝2ZΣ∫x Fr (x) dx ⋯⋯⋯⋯⋯⋯⋯⋯(3)j=1

3.1.3 Elastic hysteresis lossWhen elastic bodies repeatedly come into mutual

contact, it is known that the relation between theresultant stress and strain can be plotted as ahysteresis loop. We determined the elastic hysteresisloss 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:

2P 1 E'LP 1/2

w0＝―― dy＝（―― ――――） dy……………(4)πb 2π R

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

P dδ P 1.4

dφ＝∫w0 ―― dP＝0.65 ――――――― dy⋯(5)0 dP R 0.5 E' 0.4 L0.3

measured in the opposite direction (for convenience,the estimated stress was determined with a simplifiedsin2μmethod). The greater this difference, the greaterthe plastic flow occurring on the surface.

The residual stress on the inner edge of the rollingmark on the straight roller bearing was approximately160 MPa greater than that of the crowned rollerbearing.

From these facts, we believe that the relative slipbetween the rollers and the race led to heatgeneration and plastic fluidity at the inner edge of therolling mark on the race, resulting in the surface-initiated flaking.

3. Analysis of frictional torque3)

The amount of heat generation can be given as theproduct of frictional torque and bearing speed. Giventhis, 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 velocitydistribution 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 frictioncoefficient is μ(x), the bearing torque Ms caused byspin can be described in the form of the expression(1).

2Z n

Ms＝――Σ∫μ(x) w(x) v(x) dx ⋯⋯⋯⋯(1)ωj=1

-44-


Roller peripheralspeed

Race peripheralspeed

Slip onraceRoller

Race

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

-45-

The rolling contact distance lrd per time unit wherea roller remains in contact with a race is:

ω drlrd＝―――― …………………………………(6)

2 2

Therefore, the elastic compression workload EE is:

φωdrEE＝φlrd＝――― ……………………………(7)

4

Then, if the elastic hysteresis loss coefficient istaken as βE, the relation between the heat generationand bearing frictional torque resulting from the elastichysteresis loss is:

n

ΦE＝2ZβEΣEE＝ωME ⋯⋯⋯⋯⋯⋯⋯⋯(8)j=1

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

2ZβEΣj=1

EE

ME＝―――――― ⋯⋯⋯⋯⋯⋯⋯⋯⋯⋯⋯(9)ω

3.2 Comparison between measured frictionaltorques and calculated results

Table 2 summarizes the test conditions for thefrictional torque measurement test as well as themajor technical data for the bearings tested withcrowned rollers. The temperature was measured atthe rear face of the race. The test rig used isschematically illustrated in Fig. 12.

In both experimental and calculated results with aVG2 lubricant, plotted in Fig. 13, the frictional torquedecreases as the bearing speed increases and thefrictional torque attained was lower with the single rowbearing than with the double row bearing. Theresultant oil film parameter fell in a range of Λ=0.7 to1.6.

2b

L

dy

Contact area

Fig. 11 Roller/race contact part

Bearing tested

Load

Driving motor

Air staticpressure table

Load cell(frictional forcedetection)

Air cylinder

Fig. 12 NTN vertical type bearing frictional torque test rig

00 500 1000 1500 2000 2500 3000 3500

0.1

0.2

0.3

0.4

0.5

Fric

tiona

l tor

que

Nm

Bearing speed min-1

Experiment (single row) Experiment (double row)

Calculation (single row) Calculation (double row)

Fig. 13 Test and calculation results of bearing frictionaltorque with ISO VG 2 oil

Table 2 Frictional torque test condition

Test rig

Load

Bearing speed

Lubricant

Lubrication method

NTN vertical torque test rig

3.0 kN

1000～3000 min-1

Spindle oil ISO VG2 (31˚C)

Turbine oil ISO VG32 (37˚C)

1mR coating

(a) Operating conditions for friction torque measurement

(b) Main data for frictional torque measurement test bearing

φ60×φ 85×1 mm

φ3×7.8 mm

Race

φ3×3.8 mm

Crowned roller（φ3×7.8）

Crowned roller（φ3×3.8）

Single rowrollers

Double rowrollers


When Λ=9, the major frictional torque factor isrolling viscous resistance, which accounts for 80% ormore of the total frictional torque. Spin does notgreatly affect rolling viscous resistance. As a result,there is virtually no difference between a single rowbearing and a double row bearing, both having thesame effective contact length within a bearing.

When Λ=1, the lubrication condition is mixedlubrication and the solid-to-solid contact areaincreases. As a result, spin becomes the majorfrictional torque factor. Consequently, compared witha single row bearing, the torque is lower on a doublerow bearing, which has low spin.

The estimated temperatures on raceway surfacesobtained through FEM are summarized in Fig. 16. Thetemperature on the inner edge is higher than on theouter edge and the possibility of failure is higher withthe inner edge. In addition, it is apparent that the useof crowned rollers makes the temperature lower.

In summary, we have verified theoretically thattorque on a thrust needle roller bearing in mixedlubrication conditions can be decreased by introducingdouble row crowned rollers. A lower torque in turnhelps lower the temperature on the contact surfaceand prevents bearing material failure.

The results obtained with VG32 are summarized inFig. 14. The frictional torque remained unchanged orincreased in the experiment, while the frictionaltorque increased in the calculation. There was noapparent difference between the single row bearingand the double row bearing in either the experiment orthe 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 wasnot considered in our calculations. In addition, theresults from VG2 shown in Fig. 13 could not correctlyreproduce the parameter of solid-to-solid contact.

-46-


3.3 Analysis and discussion of torque factorsThe breakdowns of frictional torque, with a larger oil

film parameter Λ and a smaller Λ are summarized inFig. 15.

00 500 1000 1500 2000 2500 3000 3500

0.1

0.2

0.3

0.4

0.5

Fric

tiona

l tor

que

Nm

Bearing speed min-1

Experiment (single row) Experiment (double row)

Calculation (single row) Calculation (double row)

Fig. 14 Test and calculation results of bearing frictionaltorque with ISO VG 32 oil

40 0 2 1284 6 1410

80

60

100

120

140

160

Distance from bore or race mm

Straight rollerCrowned roller

Tem

pera

ture

˚C

Effective rollercontact area

Fig. 16 Estimated temperature on raceway by FEM

(a)　= 9 (b)　= 1

Fric

tiona

l tor

que

Nm

Fric

tiona

l tor

que

Nm

Single row Double row Single row Double row0

0.1

0.2

0.3

0.4

0.5

0

0.1

0.2

0.3

0.4

0.5Elastic hysteresis loss

Spin

Rolling viscous resistance

Elastic hysteresis loss

Spin

Rolling viscous resistance

Λ Λ

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

-47-

Hiroki FUJIWARA

Technical Research Dept.Research & Development Center

Photos of authors

Kenji TAMADA

Intellectual PropertyPlanning Dept.

4. Conclusion

The authors experimentally and theoreticallyinvestigated the failure mechanism of thrust needleroller bearings in low viscosity lubrication conditions.From the results of various tests, we have verified thatheat generation on the inner edge of the racewayrolling mark and the resultant plastic fluidity of theraceway surface cause surface-initiated flaking onthese bearings. We have learned that under a lowviscosity lubrication condition, spin is largelyresponsible for the frictional torque on a thrust needleroller bearing and that spin-induced slip can bedecreased and excessive heat generation inhibited byintroducing crowned rollers and a double rowarrangement. Low heat generation helps suppressplastic fluidity of the raceway, leading to a longerbearing life.

The double row roller arrangement is described indetail in the article "Double Row Thrust NeedleBearings" in this edition of the NTN Technical Review.

References1) Kenji Tamada, Factors affecting the life of thrust

needle bearings, Tribology Conference Preprints(Tokyo 2002-5), 131.

2) V.M.Hauk, Evaluation of macro- and micro-residualstresses on textured materials by X-ray, neutrondiffraction and deflection. measurements, Adv. X-rayAnal., 29(1986), 1.

3) Hiroki Fujiwara, Kenji Fujii, Torque analysis for doublerow thrust needle roller bearings, The Japan Societyfor Precision Engineering 2003 Spring MeetingScientific Lectures Proceedings, (2003), 585.

4) J.A.Greenwood and J.H.Tripp, The Contact of TwoNominally Flat Rough Surfaces, Proc. Inst. Mech.Eng., Tribol., 185, 48/71(1970-71), 625.

5) Masayoshi Muraki, Yoshitugu Kimura, Research intotraction characteristics of lubricants (2nd report)"Thermal analysis of traction with nonlinearviscoelasticity", Junkatsu, 28 (1983), 753.

6) R.S.Zhou and M.R.Hoeprich, Torque of tapered rollerbearings, Trans. ASME, J. Tribol., 113(1991), 590.


-48-


[ New Product ]

Double-Row Thrust Needle Roller Bearingsfor Automotive Air Conditioners and Automatic Transmissions

1 Introduction

In recent automotive bearing applications, smalleramounts of lower viscosity lubricants are being usedto achieve better fuel consumption. As a result, theoperating conditions for bearings have becomeincreasingly severe.

In this context, a unique problem for thrust needleroller bearings have decreased bearing life because ofroller slipping. Bearings with longer lifespans areneeded.

We recently developed "double-row thrust needleroller bearings" to cope with this market need. Thesebearings have a double-row roller arrangement toinhibit the slipping of needle rollers and to achievelonger life through prevention of surface-start failure.They also have lower torque and, through roller shapeoptimization, lower noise level.

For automotive air-conditioner compressors, wehave developed a resin cage to achieve lighter weightand lower noise. For automatic transmissions, wedesigned a new sheet steel cage with inner and outercircumferences that are secured by a full enclosuretechnique to prevent cage loosening and to enhancecage mechanical strength. These cage types aredescribed below.

2. Features of double-row thrust needleroller bearings (for automotive air-conditioner compressors)

The thrust bearings for automotive air-conditionercompressors are run with limited amounts of low-viscosity lubricant in order to improve the capability ofthe 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 increasinglyneeded, which in turn leads to decreased bearing life.

Our newly developed double-row thrust needleroller bearing (Photo 1) maintains load carryingcapacity and suppresses roller slipping. The double-row roller arrangement prevents surface-start failuresthat result from disrupted oil film on bearing rollingsurfaces. Our new bearing adopts optimal rollergeometry (optimal end face shape, roller circularityand 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.

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.

Kosuke OBAYASHI*

Double-Row Thrust Needle Roller Bearings

[Comparison with single-row bearing in termsof performance]

(NTN's test results for automotive air-conditionercompressors)1 Life (surface-starting failure) : 12 times as long2 Rotation torque : 50% reduction3 Noise : 8 dBA lower

2.1 Performance evaluationTo 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 cageand a single-row roller arrangement (Photo 2). Wesubjected both bearing types to rotation torque andrunning noise measurements after a bearing life test.This section summarizes the performance evaluationtests of both bearing types.

-49-

Photo 1 Double-row thrust needle roller bearingsfor car air conditioners

Photo 2 Current bearings

Table 1 Comparison of bearing specifications

Bearing being testedLoad

Fig 1 NTN vertical type thrust bearing test rig

Bearing being tested

Load

Fig 2 NTN horizontal type thrust bearing test rig

Bearing size

Cage type

Number of rollers

Roller length (mm)

Newly developed bearing (double-row type)

Standard bearing(single-row type)

φ60ID×φ85OD ×3W (roller diameter)

Resin cage W-type sheet steel cage

96 (48 ×2 rows) 48

3.8 7.8

2.1.1 Bearing specificationThe bearings used for the performance evaluation

have 60 mm inside diameters, 85 mm outsidediameters and are 3 mm wide (roller diameter).

Table 1 lists the data for both our newly developedbearing and the standard bearing.

2.1.2 Bearing life testFor the bearing life test, we used the NTN vertical

thrust bearing test rig illustrated in Fig. 1 and the NTNhorizontal thrust bearing test rig shown in Fig. 2.


(1) Test conditionsTwo sets of bearing life test conditions are given in

Tables 2 and 3.

-50-

5 5

5

10

203040506070809095

5 57 7 7 71101 102 103

1 12 2 23 3 3

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y, %

Standard bearingNewly developedbearing

5

5

10

203040506070809095

5 57 71 1 12 2 23 3 3

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y, %

101 102 103


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

Fig 4 Result of life test condition 2 with lean lubricationfor car air conditioners

Fig 5 NTN vertical type thrust bearing rotation torquetest rig

Table 2 Life test condition 1

(2) Test resultsThe life test result of test condition set 1 is

summarized in Fig. 3, and that from test condition set2 is given in Fig. 4.

Compared with the standard bearing, our newlydeveloped bearing exhibited a life more than 12 timesas long in terms of L10 under both test condition sets 1and 2.

2.1.3 Rotation torque measurementFor measuring the rotation torque of the bearings

being tested, the NTN vertical thrust bearing rotationtorque test rig was used.

The test rig is illustrated in Fig. 5

(1) Measuring conditionsThe rotation torque measuring conditions are listed

in Table 4.

Table 3 Life test condition 2

Table 4 Measurement condition of rotation torque

Test rig

Axial load

Bearing speed

Lubrication

Measuring duration

Number of samples

NTN vertical-type thrust torque test rig

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

1000 min-1

Category 1 spindle oil, overall coating

60 sec.

n=10, each bearing type

Rotation

Load

Bearing beingtested

Rotation torquedetection load cell

Test rig

Axial load

Bearing speed

Misalignment on race

Lubrication

Number of samples

NTN vertical thrust load test rig

P/C = 0.4 with standard bearing＊

5000 min-1

2/1000

Turbine #32, circulating lubrication

n=8, each bearing type＊

P/C : load/basic dynamic rated load

Test rig

Axial load

Bearing speed

Lubrication

Number of samples

NTN horizontal type thrust load test rig

P/C = 0.4 with standard bearing

5000 min-1

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


-51-

Fig 7 Measurement result of sound level

Fig 6 Measurement result of rotation torque

Sou

nd le

vel

dB

A

651800 min-1 3600 min-1

70

75

80

85

90

Standard bearing: W-type cage + single-row rollers

Newly developed bearing: resin cage + double-row rollers

Rot

atio

n to

rque

N・

cm

00.05 0.10 0.15 0.20 0.25

20

40

60

80

P/C



(2) Measurement resultsThe measurement results of rotation torque are

plotted in Fig. 6.These results verify that the newly developed

bearing has achieved 47 to 51% torque reductioncompared with the standard bearing.

This torque reduction was achieved by both thedouble-row roller arrangement, which reduced spinslippage, and the cage resin material.

The measured rotation torques were dependent onloading, but rotation speed caused virtually novariation.

Table 5 Measurement condition of sound level

Test rig

Axial load

Measurement distance

Bearing speed

Lubrication

Measuring duration

Number of samples

NTN sound level measuring system

100N

45˚ angle, 100ｍｍ

1800 min-1，3600 min-1

Compressor oil,droplet lubrication before running

50 sec.


2.1.4 Sound measurementFor sound measurement, the NTN sound level

measuring system was used.

(1) Measuring conditionsThe sound level measurement conditions are

summarized in Table 5.

(2) Measurement resultsThe 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.


3. Feature of double-row thrust needleroller bearings(for automatic transmissions)

Featuring a lower profile in assembled units and agreater load carrying capacity, thrust needle rollerbearings are often used for automatic transmissions inautomobiles. However, the use of lower viscosityautomatic transmission oils is increasing to alleviateshift shock and improve fuel efficiently. For thisreason, the life of bearings used in this applicationmay be shortened.

Our newly developed double-row thrust needleroller bearing (Photo 3) for automatic transmissionshas a double-row roller arrangement to inhibit rollerslipping while maintaining the necessary load carryingcapacity. The newly developed bearing featuresoptimized roller geometry (with optimized roller endface shape, roller roundness and crowning shape) toachieve a longer bearing life, lower rotation torque andquieter operation.

In addition, because thrust bearings are used ineccentric 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 fullenclosure of inner and outer ring circumferences,improving cage mechanical strength. Thisarrangement prevents separation of cage halves ordisplacement between the halves and allowsintegration of the cage with the bearing ring to form aunified structure.

[Comparison with single-row bearing](NTN's test results assuming application inautomatic transmissions)1 Life (surface-starting failure)

: Three times as long2 Rotation torque : 40% decrease3 Cage strength : 1.6 times as strong4 Sound level : Decrease of 3 dB

3.1 Performance evaluationUnder conditions that simulated automatic

transmissions, the newly developed double-row thrustneedle roller bearing was compared with a standardbearing 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 measurementsand sound level measurements. The results aresummarized below.

3.1.1 Bearing specificationsThe 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 developedbearing and the standard bearing are summarized inTable 6.

-52-


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

Double-row rollers

Full circumference enclosure

Table 6 Comparison of bearing specifications

Bearing size

Cage type

Number of rollers

Roller length (mm)

Newly developed bearing (double-row type)

Standard bearing(single-row type)

φ60ID×φ85OD ×3W (roller diameter)

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

96 (48 ×2 rows) 48

3.8 7.8

3.1.2 FEM analysisThe 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.

-53-

The maximum principle stress on the standardbearing was compared with that on the newlydeveloped bearing. As a result, the stress on thenewly developed bearing was 0.62 relative to that ofthe standard bearing rating of 1. The cage of thenewly developed bearing has increased strength.

Note: The weight of the newly developed bearing is5% lighter than the standard bearing.

Fig 8 Analytical condition

Fully constrained(inner circumference surface)Load (50% load on the outer

circumference surface and50% load on the bars andpockets)

Fig 9 Analytical result

Location where maximum principle load occurs

Newly developed bearingStandard bearing

3.1.3 Bearing life testFor 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 bearingwas three times as long as that of the standardbearing.

5 5

5

10

203040506070809095

7 71101 102

12 23

Life, h

Acc

umul

ated

failu

re p

roba

bilit

y, %


Fig 10 Result of life test for automatic transmissionsTable 7 Life test condition

Test rig

Axial load

Bearing speed

Lubrication Number of samples tested

NTN vertical type thrust test rig

P／C=0.4 on standard bearing

8000 min-1

ATF, circulating lubricationOil temperature, 100˚C

n=8 each


3.1.4 Eccentric rotation testIf 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 compressionforce in the radial direction and a shear force in therotation direction resulting from the difference betweenthe 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 addressthis problem, we checked the mechanical strength ofthe new cage shape.

-54-


Fig 12 Measurement result of rotation torque

00.05 0.10 0.15 0.20 0.25

20

40

60

80

P/C

Rot

atio

n to

rque

N・

cm



Table 10 Measurement condition of rotation torque

Test rig

Axial load

Bearing speed

Lubrication

Measuring duration

Number of samples

NTN vertical thrust torque test rig

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

1000 min-1

Category 1 spindle oil, overall coating

60 sec.


Table 9 Result of eccentricity rotation test

Fig 11 NTN vertical type thrust bearing test rigfor eccentricity rotation

LoadBearingtested

Table 8 Eccentricity rotation test condition

Test rig

Axial load

Bearing speed

Eccentricity (rotation axis)

Lubrication

Number of samples tested

NTN eccentric rotation verticalthrust test rig

P／C=0.3 on standard bearing

3000 min-1

0.5mm

ATF, circulating lubricationOil temperature, 100˚C

n=3 each

1

2

3

1

2

3

53.4

51.5

60.5

Suspended at 265.0 hrs.



↑

↑

No failure

↑

↑

BearingNo. Running time, h Remarks

Ring-shaped fractureon the cage bar bore side

Sta

ndar

dbe

arin

gNe

wly d

evelo

ped

bear

ing

For the eccentric rotation test, the NTN eccentricrotation vertical thrust test rig was used.

The test rig used is illustrated in Fig. 11.(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 inTable 9.

A ring-shaped fracture occurred on the bore side ofthe 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 ofinner and outer circumferences, exhibited improvedstrength that matches the results of FEM analysis.

3.1.5 Rotation torque measurementThe 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 aresummarized in Table 10.(2) Measurement results

The rotation torque measurement results areillustrated in Fig. 12.

As a result of the measurement, we found thatcompared with the standard bearing, the rotationaltorque of the newly developed bearing is 38-43%lower.

The primary factor contributing to the torquereduction is the reduction of spin slipping due to thedouble-row roller design.

When comparing the newly developed cage with theconventional cage, other possible contributing factorsare a smaller contact area between the cage and shaftand a smaller contact width between the cage androllers.

We also learned that the rotation torque isdependent on the load applied to the bearing testedand is least affected by the bearing running speed.

3.1.6 Sound level measurementFor sound level measurement, the NTN sound level

test system was used.(1) Measuring conditions

The sound level measuring conditions aresummarized 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 newlydeveloped bearing was approximately 3 dBA lower,and at 3600 min-1, approximately 2 dBA lower.

The factor contributing to this sound suppression isoptimization of the roller geometry.

-55-

Fig 13 Measurement result of sound level

651800 min-1 3600 min-1

70

75

80

85

90

Sou

nd le

vel

dB

A

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

Kosuke OBAYASHI

Automotive Sales HeadquratersNeedle Roller Bearings

Engineering Dept.

Photos of author

4. Conclusion

This report described our newly developed double-row thrust needle roller bearings.

In this report, the scope of application for our newlydeveloped bearings were limited to automobile air-conditioner compressors and automatic transmissionsand 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 becomeincreasingly severe. Therefore, we will remaincommitted to the development and marketing of novelbearing products that only NTN can offer, alwaysaiming to increase customer satisfaction and satisfymarket needs.

Table 11 Measurement condition of sound level

Test rig

Axial load

Measurement distance

Bearing speed

Lubrication

Measuring duration

Number of samples

NTN sound level measuring system

100N

45˚ angle, 100ｍｍ

1800 min-1，3600 min-1

ATF,droplet lubrication before operation

50 sec.



-56-



Development of In-Wheel Type Integrated Motor Axle Unit

1. Introduction

Recently, fuel-cell electric vehicles (FCEV) andelectric vehicles (EV) are being regarded as promisingnext generation automobiles because of their reducedimpacts on the environment. The drive systems forEVs, including FCEVs, can be categorized into twotypes. One type consists of a motor situated on thechassis that transmits driving power to the tiresthrough a drive train, like a conventional internalcombustion engine. The other type is an in-wheelmotor system in which a motor is incorporated intoeach wheel. On the positive side, an in-wheel motorsystem allows the provision of room space equivalentto 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 ridingcomfort. Therefore, this issue needs to be resolved sothat EVs with in-wheel motor axle units can be morecommonly used. NTN has been developing an in-wheel type integrated motor axle unit, aiming toreduce unsprung mass by creating a compacter,lighter axle unit. This report describes the design

concept and initial characteristics of this novel axleunit.

2. Newly developed axle unit specifications

We are currently developing an in-wheel typeintegrated motor axle unit that, when built into fourwheels, has power equivalent to the 1500 cc gasolineengine of a Japanese compact car. The targetspecifications 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 builtinto a tire is shown in Fig. 2.

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

NTN has been developing the in-wheel type integrated motor axle unit for electric vehicles. The unitaccomplishes 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. Thispaper introduces the design concept and basic performance of the unit.

Minoru SUZUKI*Dawei WANG*

Table 1 Target specification of axle unit

Max. output 20kW

Max. torque 490Nm

Max. velocity 150km/h (when built into 14" tires)）

Mass Approx. 20 kg

Reducer type Cycloid reducer

Reduction ratio 1/11

Motor typeAxial gap type permanent magnetsynchronization motor

Max. motor speed 15000rpm


3. Design concept

(1) Application of reducerUsually, the size of a motor is dependent on the

maximum torque generated. Since direct driving in-wheel motor axle units do not have reducers, themotors must provide greater torque, making themotors 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 maximumtorque needed for the motor by incorporating areducer and selecting a high-speed motor so that asmaller motor could fulfill the requirements.

(2) Type of reducerWe sought a reducer type that helps satisfy the

smaller space requirement and provides a greaterreduction ratio. With a 2K-H planetary gear reducingmechanism, which is a type often used as a reducersystem, the reduction ratio per stage is in the range of1/4 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 cycloidreducing mechanisms 1-3), all of which are K-H-Vplanetary gear reducing mechanisms, such as the oneillustrated in Fig. 3. With this mechanism, the finalreduction ratio is dependent on the difference in thenumbers of teeth between the gears. Thus, even onegear train can achieve a greater reduction ratio.

-57-

Fig. 1 Schematic of axle unit

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

MotorReducer

Counterweight

Needle roller bearing

Tube forming

Hub bearing

Wheel

Axle unit

Fig. 2 Installation of the unit

Fig. 4 Basic structure of reducer

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

A

B

C

S

Outside pin (internal gear)

Inside pin (output shaft)

External gear

Input shaft


-58-

0

100

70

80

90

2 4 6 8 10Input torque [Nm]

Input rotational speed=3000rpm

Rolling contact typeSliding contact type

Effi

cien

cy o

f red

ucer

[%]

Fig. 5 Efficiency of reducer

Use of this gearing arrangement allows many teethto 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 pinsare secured, and the flange member and the hubbearing inner ring member with the hub flange arefastened together using NTN's propriety tube formingtechnique, which is also applied to our 4th generationhub bearings. The tube forming technique has lead toa simple fastener structure between the hub and thereducer, allowing further reduction of the axle unitsize.

(3) Higher reducer efficiency 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 reducingmechanism is said to be lower. One possible reasonfor this is that when the eccentrically oscillatingexternal gear rotates, sliding contact occurs betweenthe 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 contactbearings into these sliding contact areas.

(4) Higher reducer speed With a cycloid reducing mechanism, in principle, the

external gear eccentrically oscillates and the inertialforce induced by this motion causes the vibration ofthe component perpendicular to the rotation axis toincrease. Usually, therefore, another set of reducers(external gear) acting in reverse phase is incorporatedto dampen this vibration. However, the couple of forceresulting from unbalanced inertia on the axes of thetwo external gears perpendicular to the rotation axisstill remains, and the effect of the unbalance isprominent when the reducers run at higher speed. Ouraxle unit incorporates a counterweight to dampen thevibration induced by this couple of force resulting fromunbalance inertia.

(5) Bearing layoutThe motor rotating shaft has a cantilever support

structure from the reducer housing. Though not shownin Fig. 1, the knuckle is formed around the reduceroutside 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 thehub bearing to the motor bearing can be lubricated

with the same lubricant, the number of seals can bedecreased, 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 absolutelynecessary for an axle unit, is located in the right endof the motor.

4. Performance Test Results

The results of evaluation tests are described belowfor the basic performance of a motor-less axle unitthat consists of a reducer and a hub bearing.

(1) The efficiency of two reducer types was measured.One reducer type was a rolling contact bearingtype that incorporated a needle roller bearing intothe sliding contact area between each inside pinand outside pin set on the reducer. The otherreducer type was a sliding contact bearing typelacking incorporated needle roller bearings. Asshown in Fig. 5, at the input rotation speed of 3000rpm, 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 rollingcontact bearing. This effect is nearly equivalent tothe effect obtained from a two-stage configurationof an involute gear type 2K-H planetary gearreducer. An oil bath method was used forlubrication.

-59-

0

5

0

1

2

4

3

5000 10000 15000Input rotational speed [rpm]

axialradial

Acc

eler

atio

n [m

/s2 ]

Fig. 6 Vibration of reducer

Minoru SUZUKI


New Product Development Dept.

Photos of authors

Dawei WANG



(2) Acceleration pickup devices were installed to theouter periphery and end face of the reducerhousing to measure the radial vibration and axialvibration on the running reducer. Themeasurement results are plotted in Fig. 6. Thoughno load was applied to the reducer, it ran smoothlywithout developing significant vibration at speedsup to 15000 rpm.


5. Conclusion

Seeking to realize a compacter and lighter in-wheelintegrated motor axle unit, we adopted a cycloidreducer 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 anactual load, the reducer can run smoothly withoutdeveloping significant vibration at speeds up to 15000rpm.

Since axle units are critical automotive components,we are going to improve the reliability of our unitthrough evaluation with various tests under actualautomobile operating conditions.

References(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 ofreducers, Japan Institute of Plant Maintenance, p22-26 (1994).

(3) Sumitomo Heavy Industries, Ltd.: Catalog C2001-5.

-60-



Development of the Mono Ring CVT- Principle and Mesurement of Transmission Efficiency -

1. Introduction

Various technologies have been invented forautomotive continuously variable transmissions(CVT).1) 2) Recently, multi-stage designs have beenincreasingly 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, CVTsrealize higher efficiency, lower fuel consumption andextremely smooth drivability.

This report introduces a compact CVT we arecurrently developing that consists of a smaller numberof components and does not need an external controlsystem.

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 employsV-pulleys, our CVT uses them in the drive section onlyand one toothed ring is inserted between the V-pulleysto 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.

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

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 automotiveapplications by changing output ratios with no actual gear steps. However, the mechanism of a CVTcan 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 beennamed 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 aprototype MR-CVT were also measured confirming the validity of the basic design.

Yuichi ITOH*Tomoya SAKAGUCHI*

Development of the Mono Ring CVT

2. Mono-Ring CVT (MR-CVT)

The internal structure of our prototype MR-CVT isillustrated in Fig. 1. The input shaft and the outputshaft 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-pulleysthat rotate concentrically and in synchronicity with theinput shaft, so the V-pulleys can shift in the axialdirection. The input shaft has ball splines to allow theV-pulleys to shift in the axial direction and transmittorque 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-pulleyscan be shifted axially.

-61-

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 twoguide 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 thecounterclockwise direction only by means of a self-adjustment 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 inputtorque is transmitted to the toothed ring by the V-

Guide roller

Guide roller

Guide roller

Guide roller

Toothed roller Toothed roller

Angular contactball bearing

Ballscrew

V-pulley

V-pulley

Ball spline

Output shaft

Output gear Output gear

Input shaft

100mm

Swing

Decelerationside

Accelerationside

Fig. 1 Schematic view of MR-CVT

Guide roller

Guide roller

V-pulley

V-pulley

Toothed roller

Toothed roller

Gear pitchcircle

Gear pitch circle

2θ

Q

Q

F1

Rt

Rg

0g

F3

0IN

0OUT

Contact point betweentoothed ring and V-pulleys

Fig. 2 Basic mechanism of MR-CVT


-62-

pulleys, the toothed ring is subjected to the frictionalforce 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. Therelation 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 ofcalculation, we took the pressure angles as zero.

RgF3=F1×―― ………………………………（1）

Rt

The force F3 also functions to force the toothed ringinto 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.

F1μ=―― ……………………………………（3）

2Q

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

Rtsinθ=μ×―― ……………………………（4）

Rg

Since the toothed ring is stable when the aboverelationship is valid, a self-adjustment mechanism isformed in which the friction coefficient μis determinedfrom the slope angle θ of the V-pulleys even ifoperating conditions vary.

This is because F1 and F3 vary according to theinput torque, so the contact force Q occurring betweenthe V-pulleys and the toothed ring automaticallyvaries.

As shown in Fig. 2, when Og-Oout and the Og-Oinare 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 outputshaft, the Og-Oout and the Og-Oin are no longerperpendicular as shown in Fig. 3. Since the contactpoint between the toothed ring and the V-pulleys shiftsto 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:

F5=F3＋F4

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

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).

F2 F2

F3F3

F5

Rt

Rg

0g

0in

0in

0out0out

0g

F4

F5

F4

F1

F1

ψ

Fig. 3 Speed-up condition Fig. 4 Slow-down condition

-63-

F5 F3＋F1 sinψ F1（Rt/Rg＋sinψ）Q= ――――――― = ――――――― = ――――――――――

2sinθcosψ 2sinθcosψ 2sinθcosψ

…………………………………………………（5）

If formula (3) is substituted into the formula (5), thenthe optimal V-pulley inclination θ relative to ψ can bedetermined as a function of ψ by using formula (6)below.

μ（Rt/Rg＋sinψ）θ= sin-1［――――――――――］ ………………（6）

cosψ

In other words, the MR-CVT can have optimalefficiency if the slope angle θ, which is determined byψ, the friction coefficient μ and the geometry of thetoothed ring when it swings on the output shaft satisfyformula (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 bythe cosine theorem. In other words, the ψ can bedetermined based on the CVT design data.

The relation between θ and r in our prototype MR-CVT 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 alsoincreases. Incidentally, we designed the subcurvatureof the toothed ring so that the contact surfacepressure between it and the V-pulleys stays below themaximum allowable level when the input torque is atmaximum.

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

Table 1 Test condition

Fig. 6 Test rig for measuring transmission efficiency

3. Test method

To verify the feasibility of the operating principle ofour MR-CVT, we developed a test rig and measuredthe 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 wereprovided with a torque meter to measure the rotationspeed and torque. For performance measurement ofthe MR-CVT, the transmission efficiency wasdetermined based on the workload, which is theproduct of the rotation speed and torque of the inputand 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 loadmotor was used for the drive motor.

The range of the gear ratio, the ratio of the rotationspeed of the input shaft to that of the output shaft, was0.3 to 1.3. The gear tooth ratio of the toothed ring (88teeth) to the output gear (60 teeth) was 1.47. The testconditions 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.

10

8

6

4

2

05 15 25 35 45 55

Radius of contact point r, mm

Slo

pe a

ngle

θ,

deg

rees

Torque meter

Torque meterMR-CVT

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

Input rotation speed, min-1

Input torque, N-m

Transmission gear ratio (radius, mm)

Dynamic viscosity of lubricant

Lubricant supply temperature


4. Test results

The measurement results for the input torque andtransmission efficiency in an input rotation speedrange 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 greaterinput 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 maximumtransmission efficiency reached 92%. Though thetransmission efficiency at the gear ratio of 0.4 waslow, this gear ratio is used only when an automobilestarts to run. The time when this gear ratio is usedaccounts for a smaller portion of the whole duration ofcar usage, so this low efficiency does not seem togreatly affect the overall efficiency of this MR-CVT inactual automobiles.

The reason for the increased transmission efficiencywith higher gear ratios seems to be a decrease in theloss that results from spin-induced slip in the contactarea between the toothed ring and the V-pulleys. Theloss of input torque at 10 N-m seems to result from agreater ratio of various frictional losses to transmissionefficiency in the MR-CVT.

-64-


100

95

90

85

80

75

700 10 20 30 40 50 60 70

Tran

smis

sion

effi

cien

cy,

%

Input torque, N・m

Gear ratio 0.4

Gear ratio 0.6

Gear ratio 0.8

Gear ratio 1.0

Gear ratio 1.2

Fig. 7 Input torque and transmission efficiency(1000min-1)

100

95

90

85

80

75

700 20 40 60

Tran

smis

sion

effi

cien

cy,

%

Input torque, N・m

Gear ratio 0.4

Gear ratio 0.6

Gear ratio 0.8

Gear ratio 1.0

Gear ratio 1.2


100

95

90

85

80

75

700 20 40 60

Tran

smis

sion

effi

cien

cy,

%

Input torque, N・m

Gear ratio 0.4

Gear ratio 0.6

Gear ratio 0.8

Gear ratio 1.0

Gear ratio 1.2


-65-

5. Conclusion

We have invented a simple, compact MR-CVT onwhich a single (mono) toothed ring swings to changetransmission speed. The MR-CVT features a self-adjustment 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 itsself-contained adjustment capability.

(2) The transmission efficiency of our MR-CVT wasat least 85% with a gear ratio of 1.0 or higherover nearly the entire speed range and themaximum transmission efficiency reached92%.

References1) 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).

Yuichi ITOH



Photo of authors

Tomoya SAKAGUCHI




-66-


[ New Product ]

Hybrid Grease NA204Ffor Automotive Electrical Instruments and Auxiliary Device

1. Introduction

To cope with the recent trend toward higherperformance 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 automotivebearings have become increasingly severe. Thegreases for bearings on automotive electricalinstruments and auxiliary devices often consist of ureacompounds as thickeners and synthetic oils as baseoils. The greases used for applications in ambienttemperatures of 180˚C or higher sometimes usefluorine compounds as thickeners and base oils.Greases containing fluorine compounds are muchmore expensive than urea greases. For this reason,strong demand exists for a less expensive, high-temperature long-life grease.

To address this problem, NTN has developedHybrid Grease, which features excellent cost-performance and high functionality. This paperdescribes the features of the Hybrid Grease and basiccharacteristics associated with its application inbearings for automotive electrical instruments and

auxiliary devices.

2. Hybrid Grease defined

Grease prices are proportional to high-temperaturelife values as summarized in Fig. 1. In other words,greases that are more expensive also boast longerlives. With the goal of ending this commonly acceptedrelationship, we have been committed to developing aless expensive, high-temperature, long life grease forautomotive electrical instruments and auxiliary

**New Product Development R&D Center Mechatronics Research Dept.**Automotive Sales Headquarters Automotive Engineering Dept.

This report introduces newly developed Hybrid Grease NA204F for automotive electricalinstruments and auxiliary devices. NA204F has a unique composition that is a mixture of ureagrease and fluorinated grease, and shows longer grease life at high operating temperature (up to200˚C) at lower cost than that of typical fluorinated grease. Moreover, NA204 is greatly improved onweak 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.

Masaki EGAMI****Mitsunari ASAO****Tomoaki GOTO****

Fig. 1 Cost and high temperature life of greases

7000

6000

5000

4000

3000

2000

1000

0 5000 10000 15000 20000 25000 300000

Price, yen/kg

180˚

C g

reas

e lif

e,

h

Lithium soap/ester

Urea/aromatic ester

Fluorine grease

Urea/PAO, urea/ester

Urea/ether

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, mixingdissimilar greases needs to be avoided since it oftenleads to changes in grease properties such asconsistency and reduced grease life. 1) However, ourattempt to mix urea and fluorine greases was anexception because we have attained extremelyfavorable grease properties.

Fig. 2 illustrates the trend in grease life with variousmixture ratios of urea and fluorine greases. The testconditions were in accordance with ASTM D 3336. A6204 bearing with a shield plate, whose internal spacewas 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 untilbearing seizure. As can be seen in Fig. 2, the life ofurea 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 assummarized in Fig. 4. At 200˚C, urea greasedeteriorates in a short time span, and reaches the endof its life in approximately 200 hours.

-67-

Cartridge heater

Test bearing

Support bearing

Pulley

Thermocouple

Loading coil spring

Test bearing : 6204 (w/ shield plate)Amount applied : 38% of internal spaceTemperature : 200°C (bearing outer ring end face)Inner ring speed : 10000 min-1

Load : Fa=Fr=67 N

Fig. 3 Grease life test rig

Before operationStart of operation→Deterioration and evaporationof urea grease→Seizure and end of life

Progress of grease deteriorationin rolling area→Reaction with rollingelements and rolling surfaces→Progress of wear

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

※Grease channeled to the sides does not contribute to lubrication.

※Residual grease principally consists of fluorine grease.

Outer ringCage

Urea grease

Start of operation→Channeling

Start of operation→Deterioration of ureagrease

Fluorine grease

HybridGrease

Inner ring

Shield plate

Rolling elements (balls)Appliedgrease

Seizure andend of life

※Full utilization of fluorine grease

Corrosion-induced wear

Film formation is a self-repairing feature

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

6000

5000

4000

3000

2000

1000

0 20 40 60 80 1000

(Fluorine 100%)(Urea 100%)

Fluorine grease compounding ratio, VOL %

Gre

ase

life,

h（

200˚

C）

Fig. 2 High temperature life of Hybrid Grease


-68-

Immediately after the start of bearing rotation, mostof the fluorine grease is shifted to the sides of theraceway, and only a small amount of grease remainson 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 remainingamount 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 membersprogresses due to the chemical reaction between thesteel surfaces and PFPE 3) 4) 5), and the bearingapproaches seizure and the end of life.

However, as the urea grease blended in our HybridGrease degrades, a film is formed on the rollingelements and rolling surfaces that inhibits the reactionbetween the steel parts and PFPE, preventing wearon them. Thus, the heat resistance of PFPE is fullydeveloped to help extend the high-temperature life ofthe Hybrid Grease.

3. Application to automotive electricalinstruments and auxiliary devices

Hybrid Grease compositions can have variousfluorine grease mixture ratios as shown in Fig. 2. Wedetermined the mixture ratio that leads to the longestgrease life. Beginning with grease with this ratio as thebase, we blended optimal amounts of additives toimprove characteristics including low temperatureperformance and rust preventive performance inaddition to high-temperature life, resulting in thedevelopment of Hybrid Grease NA204F forautomotive electrical instruments and auxiliary devices.

3.1 General properties and high-temperature lifeThe general properties of our newly developed

grease are summarized in Table 1. For comparison,this table also provides the properties of a high-temperature, long life fluorine grease A that has beenperforming well for automotive electrical instrumentsand auxiliary devices and of fluorine grease B, whichfeatures excellent cold temperature performance.

Compared with fluorine grease A, Hybrid GreaseNA204F features a high-temperature lifeapproximately 1.5 times as long, as well as a lowerlow-temperature torque. Compared with fluorinegrease B, our Hybrid Grease boasts a high-temperature life approximately six times as long,though its low temperature torque is somewhatinferior.

3.2 Cold temperature noise characteristicsDepending on the pulley specifications and

operating conditions, a belt pulley incorporating adeep-groove ball bearing can develop coldtemperature noise known as hoot noise when startedin a cold climate. With a frequency range of 500 to1000 Hz and a sound pressure that can exceed 100dBA, this noise is very offensive.

The cause of hoot noise seems to be that the pulleysystem develops resonance due to self-excitedvibration 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 rigillustrated in Fig. 5 that was at a room temperature.

Table 1 General properties of Hybrid Grease

Characteristics Unit Fluorine grease A Fluorine grease B Test methodHybrid GreaseNA204F

Diurea compoundPTFE

PTFE PTFE

PFPE PFPE

－

－

－

JIS2220.23

JIS2220.7

JIS2220.8

JIS2220.11

JIS2220.18

Per ASTM D3336

160

－

90

－

4.8 4.7

1.85 1.94

37.6 4.9

28.0 2.6

4050 965

Category 1（310～340）



more than 250

0.0

1.22

16.9

7.0

6000

Special ester oilPFPEPAO

148

－

－

－

˚C

mass %

N・m

h

mm2/s

Thickener

Base oil

60-times worked penetration

Dropping point

Oil separation (100˚C×24 h)

Specific gravity

Low temperature torque Starting(-30˚C) Steady

Grease life＊

＊Evaluation test shown in Fig. 3

Base oil dynamic viscosity (@40°C)

Start

Running

-69-

Idler pulley

Timing beltDriving pulley

Test pulley

Fig. 5 Tension pulley type test rig 60

50

40

30

20

10

0HybridGreaseNA204F

Fluorine greaseA

Rus

t occ

urre

nce

%

Fluorine greaseB

Fig. 6 Rust prevention ability of greases

Belt tension was applied and operation was startedwhen the bearing temperature rose to -20˚C and wasmaintained until the bearing temperature reached10˚C. During this period, the noise level (coldtemperature noise) was checked by listening. Foreach 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 temperatureperformance, our Hybrid Grease NA204F did notdevelop cold temperature noise and was found toexhibit excellent cold temperature noise preventionability.

Table 2 Test condition of hoot noise at low temperature

Test bearing

Radial clearance, μm

Amount of applied grease

Outer ring speed, min-1

Pulley load, N

Test temperature, ˚C

Number of repetitions

6203 (light-contact type, w/ rubber seal)

12～16

38% of internal space

2700

127

－20～10

n=2×5 times

Table 3 Rate of hoot noise generation at low temperature

Hybrid GreaseNA204F

Fluorine grease A

Fluorine grease B

60

0

0

○

○

○

○

○

○

○

○

○

△

○

○

○

○

×

×

○

○

○

○

×

×

○

○

○

○

×

×

○

○

No.1

1 2 3 4 5

No.2

No.1

No.2

No.1

No.2

GreaseItem Number of repetitions Occurrence

％

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

3.3 Rust prevention abilitySince ordinary rust preventing grease additives do

not dissolve in PFPE, they cannot be used in fluorinegreases. 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 haspowerful rust prevention ability.

Fig. 6 summarizes rust occurrence percentagesobtained from an evaluation test conducted accordingto ASTM D1743 in which salt water was used insteadof pure water. For evaluation, a 4T-30204 taperedroller bearing coated with each of the test greaseswas immersed in 1% aqueous sodium chloridesolution for 10 seconds, and was allowed to stand for48 hours in an atmosphere of 40｡C and 95% relativehumidity. 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 fluorinegreases A and B, Hybrid Grease NA204F exhibitedexcellent rust prevention ability.


4. Compatibility with seal rubber material

The rubber seal material in bearings with fluorinegrease is usually a fluorine-containing rubber in orderto satisfy heat-resistance requirements. The mostcommonly used fluorine-containing rubber materialsare so-called FKMs (an ASTMD1418 abbreviationrepresenting fluorine-containing rubber) that include abinary copolymer of vinylidene fluoride andhexafluoropropylene (VDF-HFP) and a ternarycopolymer comprising VDF-HFP andtetrafluoroethylene (VDF-HFP-TFE). These commonFKMs 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 fluoridecomponent 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 eventuallylost. 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 anti-urea 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 testpieces made of NUF and common FKM that wereimmersed in Hybrid Grease NA204F at 200˚C. Whilethe physical properties of the common FKM greatlydegraded 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 withHybrid Grease, but also when it is used together withurea grease.

5. Conclusion

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 longerthan that of conventional fluorine greases. At thesame time, it is free from the drawbacks associatedwith fluorine greases, including poor rust preventionability and a tendency toward leakage induced by rustpreventive oil within bearings. Furthermore, it helpsprevent occurrence of cold temperature noise whenapplied to pulley bearings. Being less expensive thanfluorine greases, Hybrid Grease NA204F boastsextremely high cost performance, and when usedtogether with our newly developed urea-resistant

3.4 Leakage characteristics of greasesFluorine grease is not compatible with the base oil

(mineral oil) in rust-preventive oil that is used in thebearing manufacturing process. As a result, if fluorinegrease is inserted in a bearing coated with such rust-preventive 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 greaseapplication in bearings, it is common to limit theamount of residual rust preventive oil within thebearing.

By following a manufacturing practice for applyingcommon 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 10minutes. The resultant grease leakage percentages(expression 1) are summarized in Fig. 7.

amount of leakageLeakage percentage＝―――――――――――――×100 (%)

original inserted amount

Expression (1)

Large amounts of fluorine greases A and B leakedfrom the internally rustproofed bearing. In contrast,only a small amount of Hybrid Grease NA204Fleaked from the internally rustproofed bearing.

As mentioned in Sec. 3.3, with its powerful rust-prevention ability, Hybrid Grease will help contribute tobearings that are more resistant to rust compared tofluorine grease-filled bearings.

-70-


Fig. 7 Leak rates of greases during rotation

Test bearing : 6204 (w/ non-contact rubber seal)Applied amount : 38% of internal spaceInner ring speed : 10000 min-1

Load : Fa=39 NTemperature : room temperatureOperating time : 10 min.

12

10

8

6

4

2

0HybridGreaseNA204F

Fluorine greaseA

Gre

ase

leak

age

%

Fluorine greaseB

Degreased bearingInternally rust-proofed bearing

-71-

fluorine rubber, NUF, it does not developincompatibility problems with seal rubber material.Thus, we believe that Hybrid Grease NA204F willgreatly contribute to performance enhancement andcost reduction for automotive electrical instrumentsand auxiliary devices.

References1) Tadashi Hatakeyama: Mixing of dissimilar greases,

and variation in properties, Junkatsu, 19, 4 (1974)331.

Masaki EGAMI



Photos of authors

Mitsunari ASAO



Tomoaki GOTO


Hardening

of surface

Deformation

Hardening

of surface

Deformation

NUFCommonFKM

CommonFKM

NUF

Immersion time, h

Var

iatio

n, %

Var

iatio

n, %

Immersion time, h 200˚C, in NA204F

200˚C, in NA204F

(a) Hardness

(b) Tensile Strength

00

0

5

10

15

20

-5

-10

100

100 200 300 400 500 600 700 800 900 1000

200 300 400 500 600 700 800 900

-20

-40

-60

-80

-100

10000

Fig. 8 Deterioration of fluorinated rubbers dipped into Hybrid Grease at 200˚C

2) Japanese Society of Tribologists: TribologyHandbook, 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-basedlubricants, Tribologist, 42, 3 (1997) 205.

5) Mineo Suzuki, Philippe Plat, Koji Matsumoto: JAXAresearch and development report, Performance ofball bearings having a retainer impregnated withspace lubricants and their outgassing properties,JAXA-RR-03-014 (2004).


-72-


[ New Product ]

Ballscrew for Automated Manual Transmission

1. Introduction

Recently, an increasing number of automobilesmodels have been incorporating automated manualtransmissions (AMT) to improve fuel efficiency andride comfort.

In the past, most automated gear change systemsfor 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 arecentered on worm gears or sliding screws. theseattempts had drawbacks including higher cost due tothe increased number of components, inability to offergreater driving power due to relatively large friction,lower efficiency, and poor response.

To solve these drawbacks, Mitsubishi Fuso Truckand Bus Corporation and Bosch Corporationdeveloped a gearshift actuator for a compact truck

automated manual transmission (INOMAT-@) inwhich 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.

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 achievedwithout significant change to the transmissionstructure. The significantly altered area is shownencircled by a dashed line.


Recently , the use of automated manual transmissions (AMT) are expanding in the automotivemarket. Of the many methods of automation in the market, the ball screw and driven by DC motorare 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 bearingsfor this application.

This article documents the development and future applications.

Koji TATEISHI**Keisuke KAZUNO**


2.2 Overall layout of gearshift actuatorFig. 1 illustrates the general structure of the

gearshift actuator.1)

The gearshift actuator achieves gear change bysynchronously driving the two ball screwsperpendicular with each other by means of DCmotors. The gearshift actuator is capable of rapidresponse 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.

-73-

Photo 1 Transmission

Ball screw

Ball screw

Non-contactsensor

Non-contactsensor

Select motor

Sliding lever Shifting shaft

Shift motor

Fig. 1 Construction of gear shift actuator

Photo 2 Ball screw for AMT

3. Features of ATM ball screws

The ball screws used in this ATM are shown inPhoto 2.

3.1 Ball screw circulation sections The circulation section on each of the ball screws is

a deflector type. Thanks to a new staking methodeach deflector Photo 3. is very compact with a radialdimension approximately 20% smaller than ordinaryreturn tubes.

The deflector staking assembly is shown in Photo4. 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, butis reliable because it has sufficient strength againstloosening of the deflector member toward the outsideof the nut. The deflector is formed with high-precisionmetal injection molding (MIM), guaranteeing stableball circulation.

Photo 3New deflector

Wing

Photo 4Appearance of deflector

staking assembly

Photo 5Appearance of deflector staking

For shift side

For select side

Gear shiftactuator

Transmissionbody


-74-

3.2 DurabilityWhen the gearshift actuator is running, the fork-

shaped 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 onthe ball screw from the trunnion because of factorsincluding 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 longerball screw life, improving the load distribution on theballs is important.

To address this problem the authors have changedthe ball circulation system from a return tube systemto a deflector system that optimizes the span andcircumferential phase of the circulation deflectorrelative 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.

Table 1 Specification of ballscrew

5000

4000

3000

2000

1000

0-15 -10 -5 0 5 10 15

Coordinate in ball axis direction (mm)

Con

tact

pre

ssur

e (M

Pa)

20% 16%

Contact pressure reduction measure

※The loading position is taken as zero.

◆ Tube type ▲ Deflector type

Fig. 3 Comparison of ball contact pressure

Characteristics Deflector typeTube type

Shaft dia. (mm)

Lead (mm)

Ball dia. (mm)

Number of circulations (rounds×rows)

Moment load (N-m)

φ14.5

4.0

φ2.778

17

1×43.5×1

Fork-shapedarm Fork-shaped

arm

Nut trunnion

Shift shaft Shift shaft

Fig. 2 Driving way of ball nut

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 onballs can be reduced by as much as 20%. Sufficientdurability was not attained with any return tubesystem, but the NTN deflector system has achievedthe necessary durability.

4. Evaluation test

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

4.1 Operability (efficiency)Shift side ball screws must be capable of

developing sufficiently large thrust because they aremeshed with a shift gear. Fig. 4 summarizes theresults of the relationship between the ball screwbacklash and the resultant thrust.

-75-

Because a smaller radial backlash on any ballscrew reduces the degree of freedom of the nut, thefriction on the sliding surface of the actuator willincrease 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 themoment load.

From the test results, we determined an optimalbacklash on the ball screw that ensures thrustnecessary for shifting action.

4.2 DurabilityThe tests and test conditions for evaluating the

durability of the ball screws are summarized in Tables2 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 thedurability test.

10000.06 0.08 0.10 0.12 0.14 0.16 0.18 0.20

16001700

1800

1900

2000

1500

14001300

12001100

y=1191.9x+1536.2

Radial backlash on ball screw mm

Thr

ust o

n ba

ll sc

rew

N

Fig. 4 Relationship between ballscrew's lash and thrust

Koji TATEISHI


Photos of authors

Keisuke KAZUNO


Table 2 Durability test condition of shift side ballscrew

Table 3 Durability test condition of select side ballscrew

Test Test conditions

Max. load

Stroke

Drive motor

Ambient temperature

Lubrication (grease)

Number of operations

Impact load

Ambient temperature



Vibration frequency

Acceleration

Ambient temperature


Duration of vibration

2800N

±5mm

24V DC motor

110˚C

Ether-based synthetic oil + aromatic diurea

2,310,000

8000N (two directions)

Room temperature


2000

25～400Hz

13.7～96.1m/s2

Room temperature


12h

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

Dur

abili

ty te

stS

tren

gth

test

Fret

ting

test

Max. load

Stroke

Drive motor

Ambient temperature

Lubrication


Vibration frequency

Acceleration

Ambient temperature

Lubrication

Duration of vibration

Inertia load by acceleration and deceleration of nut

±7mm

24 VDC motor

110˚C

Contaminated transmission oil

2,200,000

25～400Hz

3.7～96.1m/s2　*In three directions

Room temperature

Contaminated transmission oil

12h

Test Test conditions

Dur

abili

ty te

stFr

ettin

g te

st5. Conclusion

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 ballcirculation section. By applying these improvements,NTN is currently developing ball screw units for use inthe electrically driven automotive actuators of enginebrake controllers in an attempt to expand the scope ofapplications for these ball screws.

Reference1) Nakamura, et al. : Development of new INOMAT,

Mitsubishi FUSO Technical Review (2004, No. 1), p.49-54


-76-


[ New Product ]

Drawn Cup Needle Roller Bearings for Throttle Bodies

1. Introduction

A throttle valve is a device that controls the amountof air fed into an engine. To cope with recent growingneeds for better fuel efficiency and lower emissions,electronically controlled throttle valves that arecapable of fine adjustment of an air/fuel ratio are beingused 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 usedincreasingly to enable both a compact lightweightdesign and an improved response with the motor usedfor this purpose.

Rolling contact bearings are being used in thesystems that control the airflow into automotiveengines as well. The support for the throttle blademust not allow air to leak, therefore rolling contactbearings must have good sealing performance.

To address these needs, we have developed drawncup needle roller bearings for throttle bodies that haveseals with lower friction and less air leakage. Thispaper introduces the features and performance ofthese drawn cup needle roller bearings.

2. Structure and features

The rolling bearings used for electronically controlledthrottle valves can be categorized as deep grooved ballbearings (hereafter, referred to as "ball bearings") anddrawn cup needle roller bearings (hereafter, referredto as "drawn cup bearings") (Fig. 1).

Compared with ball bearings, drawn cup bearingshave smaller outside diameters, and as a result, theradial space of the housing can be smaller and thehousing can be lighter. Ball bearings are narrowerthan drawn cup bearings, but their built-in seals alonecannot completely prevent air leakage so additionalseals must be incorporated, which leads to the needfor additional space and expense.

Our newly developed drawn cup bearing assemblyconsists of a drawn cup bearing fabricated by the pre-bent technique* and a seal that does not have a coremetal. This bearing does not need additional sealing,and it can contribute to environmental protectionthrough cost reduction, space conservation andreduced weight.

*Pre-bent technique: a technique in which a cage and rollers arefit into a drawn cup outer ring, and then the assembly issubjected to heat treatment.

*Automotive Sales Headquarters Needle Roller Bearings Engineering Dept.

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.

Hideki AKAMATSU**

Drawn Cup Needle Roller Bearings for Throttle Bodies

3. Friction Performance

The friction performance of the bearings thatsupport the throttle blade is a critical factor in adoptinga compact design and achieving quick response withthe motor.

A friction measurement value is obtained byinserting a torque gauge with a shaft of a specifieddiameter into a drawn cup bearing that is press-fittedinto a standard ring, and turning the gauge 90degrees. The friction value of these bearings isapproximately 0.8 N-cm, which is sufficiently lowerthan the development target of 1.1 N-cm (Table 2)

-77-

Table 1 Specification of drawn cup needle roller bearingsfor throttle bodies

Throttle blade Return spring

Drawn cup bearing (w/ built-in seal)

Motor

Control gearing

Separate seal

Deep grooveball bearing

Throttle blade Return spring

Motor

Control gearing

Fig. 1 Structure of electronic throttle bodies

Fig. 2 Structure of drawn cup needle roller bearingsfor throttle bodies

Table 2 Test result of friction

1.1 N-cm max.

30 pcs.

0.78 N-cm

0.85 N-cm

0.70 N-cm

0.054 N-cm

Development target

Number of samples (n)

Mean value

Max. value

Min. value

Standard deviation

Deep groove ball bearing application exampleDrawn cup bearing application example

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

Chromium molybdenum steel

V-type welded cage

Crowned

Fluorine rubber, no core metal

Dynamic rated load Cr 4500 NStatic rated load C0r 5100 N

High performance grease applied

1.1 N-cm max.

0.8cm3/min. max.

Dimensions

Outer ring

Cage

Rollers

Seal

Rated load

Others

Friction

Air leakageDeve

lopm

ent

targ

et

Spe

cific

atio

ns


-78-

Suction

Suction

Negativepressure

Test bearing

Standard control

Atmosphericpressure

Photo 1 Air leakage tester

Fig. 3 Measurement part

4. Air Leakage Prevention Performance

Because the throttle valve regulates the amount ofair fed into the engine, the amount of air leakage fromthe supports of the throttle blade must be strictlycontrolled.

The measuring method and resultantmeasurements of the air leakage on drawn cupbearings are summarized below.

Before starting the air leakage measurements, eachdrawn cup bearing was press-fit into a ring of specifieddimensions, and the shaft was inserted into thebearing of a specified diameter. This set was thenmounted to a sufficiently rigid measuring head. Theother 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 theair leakage tester, and Fig. 3 illustrates themeasurement heads.

The amount of air leakage with our drawn cupbearing was 0.01 cm3/min. or lower, which issufficiently lower than the development target of 0.8cm3/min (Table 3).

Table 3 Test result of air leakage

0.01cm3/min. max.

0.01cm3/min. max.

0.01cm3/min. max.

0.01cm3/min. max.

BearingNo.

Air leakageDevelopment target: 0.8 cm3/min. max.

φ9.975

No.1

No.2

No.3

No.4

Diameter of shaftinserted (mm)

Hideki AKAMATSU

Automotive Sales HeadquratersNeedle Roller Bearings

Engineering Dept.

Photos of author

5. Conclusion

This paper has introduced our unique drawn cupbearing assembly for a throttle valve. This assemblyincorporates a compact, less expensive drawn cupbearing and a seal that features lower friction and airleakage prevention capabilities. Use of this bearinghelps reduce the size of the throttle valve and improvethe 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 globalenvironmental protection.

-80-


[ New Product ]

Compact and Shudderless Constant Velocity Joint (EPTJ)

1. Introduction

Constant velocity joints (CVJ) used for automotivedrive 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 startedor accelerated and the vibration that occurs duringidling of A/T cars are affected by vibrationcharacteristics that result from the induced thrust andslide resistance of sliding-type constant velocity joints.

In an attempt to solve these problems, in 2002, NTNdevelopes the revolutionary PTJ constant velocity jointwith vibration characteristic values that have beenreduced by more than 50% compared to conventionallow-vibration joints. Since then, the production of thisnovel constant velocity joint has been expanding.

With our PTJ products, the roller cassette in thejoint slides smoothly on the outer race roller groovewhile maintaining specific attitude even when the jointis 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 C.V.Joint Engineering Dept.

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) characteristicin vehicles. The high efficiency E-series PTJ, what we call EPTJ, is also lighter and more compactthan a conventional PTJ. The EPTJ maintains the excellent vibration characteristics of the ultra-lowvibration 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.

Tatsuro SUGIYAMA**Yuuichi ASANO**

Front wheel drive axle

result, the induced thrust from the running PTJ hasbeen reduced by more than 70% (at an operatingangle of 6˚ or greater) compared to conventionaltripod joints, greatly reducing the shudder that occurswhen a vehicle is started or accelerated.

In addition, our constant velocity joint design hasgreatly reduced sliding resistance, and significantlyimproved 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 through

Sliding-type CVJ

Fixed-type CVJ


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

2. EPTJ structure

The structure of our newly developed EPTJ isillustrated 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 (innerrace/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 (NTNdata: one size reduction) (Fig. 2). The weight of theEPTJ is approximately 8% less than that of the PTJthrough size reduction of the outer race and innercomponents.

Fig. 1 Construction of PTJ and EPTJ

Fig. 2 Outer-diameter comparison between PTJ and EPTJ

-81-

105

100

95

90

85

80

75

70

70 80 90 100 110 12065

Out

er r

ace

outs

ide

dia.

mm

NTN joint size

PTJ

PTJ

EPTJ

EPTJ

Fig. 3 Measurement of insulation resistance

EPTJ

Competitor's low vibration CVJ

00

10

20

30

40

50

60

2 4 6 8 10 12 14Indu

ced

thru

st 3

rd o

rder

com

pone

nt

N

Operating angle deg.

Conventional product: PTJ

Newly developed product: EPTJ

Outer race

Trunnion

Roller cassette×3

¡Outer race ¡Inner race ¡Sap ring×2¡Needle roller× n

▲4％

3. EPTJ functions

Making CVJs more compact and lightweight leadsto reduced vibration performance, mechanicalstrength and bearing life. We have overcome thesedrawbacks and our EPTJ has attained functionalitycomparable to the PTJ.

3.1 Reduction of induced thrust load(Countermeasure against accelerating vehicle shudder)

Induced thrust load on a CVJ is an axial forceinduced by mutual friction between joint componentswhen the shaft rotates at an angle with the outer race.This thrust is associated with shudder in anaccelerating car, so reducing it can help suppress carshudder.

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 contactrelationship, the moment of roller cassette tilt while theCVJ is running is smaller and the attitude of the outerrace 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 EPTJare nearly equivalent to those with the PTJ.


-82-

4. FEM stress analysis and mechanicalanalysis

To improve the overall strength, durability and NVHcharacteristics in good balance, we determined theoptimal geometry of the major components (outerrace, trunnion and needle rollers) through fullutilization of FEM stress analysis. We sought lowerstress levels and compared evaluation results withactual components.

We have also fully utilized dynamic mechanismanalysis, as shown in Fig. 6, to understand theinternal forces and component motions that occur inoperation of the EPTJ in order to optimize its internaldimensions.

3.2 Reduction in static plunging slide resistance(Countermeasure against D range idling

vibration with A/T cars)Slide resistance on a CVJ, which occurs when the

outer race is vibrating and allowed to slide in the axialdirection, is a cause of D range idling vibration in A/Tcars. Therefore, idling vibration can be suppressed bycontrolling the slide resistance.

Like the PTJ, the EPTJ is capable of oscillation onthe roller cassette bore and the trunnion journal. Inother words, with the EPTJ, the roller cassette rollsonly 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 needlerollers, thereby keeping the slide resistance of theEPTJ 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.

EPTJ

Competitor's low vibration CVJ

0

20

40

80

100

60

4 6 8 10 12

Sta

tic p

lung

ing

slid

e re

sist

ance

N

Angle deg.

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

（MPa）500

300

100

-100

Fig. 5 FEM stress analysis for outer-race

Roller cassette inclination

Fig. 6 Dynamic mechanism analysis

3.3. StrengthThe mechanical strengths of the internal

components of the EPTJ, including static torsionstrength and torsion fatigue strength, are greater thanthose of outer race stems and shafts, meaning thatthe internal components are strong enough for use inreal cars.

3.4 DurabilityThe 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. Inaddition, the EPTJ durability is equivalent to or betterthan that of a competitor's CVJ.

-83-

Photos of authors

5. Application of quality engineering

The EPTJ is essentially a compact version of thePTJ. Size reduction can cause problems includingdecreased component rigidity and adverse effects onbearing functions and characteristics due to theincreased contact pressure. In particular, the compactdesign significantly affected the NVH characteristics,and the induced thrust at a high torque load. Toaddress this problem, we applied quality engineeringto achieve a robust design.

In application of quality engineering, we first clarifiedthe functions of low vibration constant velocity joints,and then listed the controllable factors, operatingconditions and variations in manufacturing that couldaffect their functionality, before performing anexperiment using the L18 orthogonal matrix. Weselected the factors that significantly affect thefunctions of constant velocity joints and adopted ourfindings in designing the EPTJ.

As shown in Fig. 7, the induced thrust on the EPTJwas decreased by approximately 50% throughapplication of quality engineering.

Tatsuro SUGIYAMA

Automotive Sales HeadquartersC.V.Joint Engineering Dept.

Yuuichi ASANO


Indu

ced

thru

st 3

rd o

rder

com

pone

nt

N

Indu

ced

thru

st 3

rd o

rder

com

pone

nt

N

00 5 10 0 5 10

50

100

150

0

50

100

150

Operating angle deg. Operating angle deg.

Before application ofquality engineering

After application ofquality engineering

Load torque A

Load torque B

Load torque CLoad torque D

Load torque A

Load torque B

Load torque CLoad torque D

Fig. 7 Quality engineering for induced cyclic load reduction

6. Conclusion

EPTJ features:(1) The magnitude of induced thrust is not affected by

the operating angle of the joint, and the EPTJ canattain low vibration comparable to the PTJ.

(2) The magnitude of static plunging slide resistance isnot affected by the operating angle of the joint, andthe EPTJ can attain low vibration comparable to thePTJ.

(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 CVJsand that is minimally affected by the operating angle.EPTJ units will fully demonstrate their designperformance in conditions when better spaceefficiency is needed and a larger operating angle isrequired due to the layout limitations of a particularcar, and when greater torque is applied to the CVJwhen the car is idling, starting or accelerating.


-84-


[ New Product ]

Constant Velocity Steering Joint (CSJ)

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, ifonly one cross joint is used to provide a largeroperating angle for an automotive steering system,then variation in rotational angle and resultantvariation in torque occurs, leading to uncomfortablesteering. To avoid this problem a feature similar to thatobtained with constant velocity joints can be achievedif two joints are combined such that their operatingangle is the same. They are positioned in rotationalphases where the deviation in rotational velocity ofone 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 provideconstant velocity features.

Using our expertise with constant velocity joints thathave performed outstandingly with automotive drive

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

2. Conventional technology

Steering joints are expected to operate smoothlyand 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 velocitysteering 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 greatersize, heavier weight and low torsional rigidity. D-CJscan be categorized into two types: pseudo constantvelocity and full constant velocity. Both types havedrawbacks, including unreliable constant velocity withangles outside the preset operation angle range and anarrower oscillating angle (see Table 1).

In an automotive steering system, Cross or Cardan joints typically were used. However, because ofthe 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 jointdeveloped by NTN. Adding the original idea to the reliable know-how of constant velocity universaljoint, a smooth and non-backlash feeling has been achieved.

Kenta YAMAZAKI**



3. CSJ structure

The basic structure of the CSJ resembles that of theBJ (fixed constant velocity joint of track offset type)that is often used for drive shafts. However, with anordinary BJ, clearance between the ball groove(formed between the outer race and inner race) andthe balls is small, and as a result, backlash occurs inthe 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 clearancebetween 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 addeda plunger and a spherical plate to the joint (see Fig.1). More specifically, a plunger comprised of a springand a ball is provided at the shaft end, and the force ofthis 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, therebythe clearance on the ball groove is reduced.

With this structural arrangement, even when minorwear occurs on contact areas between thecomponents due to extended steering system use,associated clearances are automatically reducedthrough the action of the spring, maintaining thesteering system in a backlash-free state.

NTN has also attempted to optimize the shapes ofthe components and the clearance between them, andhave optimized the force of the spring accordingly. Inthis way, we have achieved both reduction inrotational backlash throughout the operating anglerange and smoother steering system rotationalmotion.

4. Performance characteristics

4.1 Rotational backlashThe 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 currentlyused double Cardan joints (D-CJ), for automotivesteering systems.

Fig. 1 Structure

Fig. 2 Structure to reduce rotational play

Inner raceBand

Outer race

Spherical plate

Ball

Shaft

BootCage

Plunger

Band

Fig. 3 Rotational back lash

Tors

iona

l tor

que

CSJD-CJ

0

0-

- +

+

Torsional angle

Spherical plate

Plunger

Displacement inaxial direction

-85-

CSJ


-86-

5. Endurance and strength

The results of an endurance test are summarized inFig. 5, and the results of strength tests are plotted inFigs. 6 and 7.

In terms of both endurance and strength, the CSJsatisfies targeted levels, and we believe that the CSJcan 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 forautomotive steering systems.

Fig. 6 Torsional fatigue strength

103±0

104 105 106 107

Number of twists

Tors

iona

l tor

que

No failure

Development target

Fig. 7 Static torsion strength

Table 1 Comparison of constant velocity steering joint

0 10 20 30 40 500

Joint angle

Failu

re to

rque D

evelopment target

Fig. 5 Endurance test results

Condition A

Test result

No problem

No problem

Operating time

Targeted life Targeted life×2

Condition B

4.2 Rotational torqueThe results of the comparison of the CSJ and the D-

CJ are summarized in Fig. 4.Though somewhat greater than that of the D-CJ, the

rotational torque of the CSJ falls in a permissiblerange and has less variation. Therefore, these CSJcharacteristics are suitable for steering joints.

Fig. 4 Rotating torque

Rot

atio

nal t

orqu

e

CSJ

D-CJ

00 60 120 180 240 300 360

+

Rotational angle deg.

Outside dia.

Weight

CSJ

φ64 mm

441ｇ

160 cm3

0～48 deg

φ61.5 mm

970ｇ

327 cm3

0～47 deg

φ70 mm

830ｇ

270 cm3

40～48 deg

Double Cardan jointfrom competitor A

Double Cardan jointfrom competitor B

(Fully constant velocity over whole range)

(Pseudo constant velocity type)※

(Fully constant velocity type)

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

Operatingangle

Volumeoccupied

-87-

Fig. 8 Crose joint (Cardan joint)

Fig. 9 Double cardan joint

7. Conclusion

Developed as a dedicated constant velocity steeringjoint for automobiles the NTN CSJ is characterized bya unique preload applying mechanism incorporatedinto the joint and internal clearance optimized torealize not only inhibition of rotary-direction backlashbut also the capability for smooth rotational operationover the entire operating angle up to the maximumoperating angle of 48 degrees. These features, whichare lacking in conventional BJs, make the CSJcapable of providing an ideal constant velocity forautomotive steering systems, while maintainingsufficient 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 rotationaldirection. In contrast, two Cardan joints are needed forthis application. Furthermore, compared with doubleCardan joint arrangements, the CSJ features lighterweight and a more compact design, allowingsignificantly 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.

Reference1) The Society of Automotive Engineers of Japan,

Automotive Engineering Handbook, Design Volume(1991)

Kenta YAMAZAKI


Photos of author

1

1 Bearing

2 Cross spider

3 Bearing seal

4 Yoke

32

4

Socket Outer sealInner sealSocket yoke

Inner seal

Bearingcup

Socket yoke

Spider

Coupling yokeBushing

Centering ballNeedle rollers

Needlerollers


-88-


[ New Product ]

Summary of E Series, Constant Velocity Jointfor Propeller Shaft

1. Introduction

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


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 serieshigh efficiency constant velocity joints for drive shafts.

This paper summarizes the E series CVJs for propeller shaft products.

Tomoshige KOBAYASHI**Teruaki FUJIO**

Table 1 Characteristics of E series

Photo 1 Example of Propeller shaft. (2 joint design)

Newseries

Maximumallowable angle

Torqueloss

Heatgeneration

Subject ofcomparison

Outsidedia. Mass

HEBJ

HEDJ

HETJ

BJ

DOJ

TJ

＋

＋

＋

＋＋＋

＋＋

＋＋

+++ …Excellent +++… Good +… Equivalent

＋＋

＋＋

＋

＋＋

＋＋

＋＋

＋＋＋

＋＋

＋＋＋

2. Overview of E Series CVJ productsfor propeller shafts

We have developed high-speed variants of the Eseries -the HEBJ, HEDJ and HETJ- whose basicstructure derives from that of other E series drive shaftproducts. Compared with conventional mass-produced CVJs for propeller shafts (BJ, DOJ and TJ),the HEBJ, HEDJ and HETJ series products boast highefficiency, light weight and compact size whilemaintaining the load carrying capacity and durability ofthe 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

Summary of E Series, Constant Velocity Joint for Propeller Shaft

chromium-free material for CVJ components that needsurface treatment.

3. E Series Propeller Shaft CVJ productfeatures and future developments

3.1 HEBJ3.1.1 Features

When a CJ (cross joint) cannot satisfy requirementsfor a greater operating angle, a BJ type CVJ forpropeller shafts is used together with a sliding typeCVJ.

Like the EBJ for drive shafts, the HEBJ for propellershafts features temperature increase and transmissionefficiency that is much improved over conventional BJproducts.

The construction of HEBJ is illustrated in Fig. 2. Asize comparison of the HEBJ and a conventional BJfor propeller shafts is provided in Fig. 2, and theiroutside diameters are shown in Table 2.

Fig. 2 Comparison of structure between BJ and HEBJ

-89-

Fig. 1 Composition of HEBJ

Outer raceCage

Ball

Inner race

Fig. 3 HEBJ flange type design

Fig. 4 HEBJ disc type design

Table 2 Comparison of outer-diameter betweenHEBJ and BJ

NTN joint size BJ HEBJ

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

87

100

79.9

91.0

72.6

83.7

Unit: mm

The flange type and disk type designs are availablefor the HEBJ in order to cope with user requirements.(See Figs. 3 and 4.)

BJ HEBJ

3.1.2 Future developmentWe 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.


-90-

3.3 HETJ3.3.1 Features

The permissible slide amount is greater with TJ typeconstant velocity joints for propeller shafts. TJproducts feature smaller slide resistance over thewhole sliding range as well as lower heat generation.

As for the ETJ for drive shafts, the internal design ofthe HETJ has been improved based on data obtainedfrom FEM analysis. As a result, the HETJ products forpropeller shafts feature a smaller outside diameter,while their strength, durability and NVH characteristicsare 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 forpropeller shafts is given in Fig. 7, and a comparison ofoutside diameters of the TJ and HETJ is given inTable 4.

3.2 HEDJ3.2.1 Features

The DOJ type CVJs for propeller shafts are capableof a greater axial slide, and can absorb minute axialvibration.

Like the EDJ products for drive shafts, thetemperature increase on the HEDJ is limited and aninduced thrust on the HEDJ is lower compared withthe conventional DOJ. The size of the HEDJ isapproximately 4% smaller than that of the DOJ. Tocope with varying CVJ use specifications for propellershafts, the HEDJ is available in sizes #71, #82, #92and #95. The standard HEDJ forms available are thepressure-welding type and flange type, both of whichare in demand.

The joint construction (pressure-welding type) isillustrated in Fig. 5, and a comparison of outsidediameters with the DOJ is given in Table 3.

Fig. 5 Composition of HEDJ

Outer race

Cage

Ball

Inner race

Fig. 6 Composition of HETJ

Outer race

Needle roller

Spherical rollersTrunnion

Table 3 Comparison of outer-diameter betweenHEDJ and DOJ

NTN joint size DOJ HEDJ

71

75

82

87

92

95

―

73

―

79

―

―

63

―

73

―

79

83

Unit: mm

Fig. 7 Comparison of structure between TJ and HETJ

TJ HETJ

-91-

Photos of authors

3.4 Future development of sliding typesAn 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).

Tomoshige KOBAYASHI


Teruaki FUJIO


Table 4 Comparison of outer-diameter betweenHETJ with TJ

Table 5 Comparison of characteristics between propeller shaft CVJ

NTN joint size TJ HETJ

75

79

68

73

66

68

Unit: mm

4. Conclusion

By applying the experience and expertiseaccumulated through the development of the E seriesCVJs for drive shafts to CVJs for propeller shafts, NTNhas successfully developed the lightweight, compact Eseries CVJs for propeller shafts.

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

Fixed type Sliding type

Slide amount

Heat generation

Induced force

Radial direction

Permissible angle

High speed operation

Rotation direction

Slide resistanceNVHcharacteri-

stics

Backlash

HEBJ HEDJ HETJ HLJ LJ

―

◎

○

○

―

8th order

―

○

◎

◎

○

○

○

8th order

○

○

○

○

◎

△

△

6 and 3th order

◎

◎

△

◎

◎

△

△

6 and 3th order

◎

◎

◎

△

△

◎

◎

3th order

△

△

◎：Excellent　　○：Good　　△：Average

Summary of E Series, Constant Velocity Joint for Propeller Shaft

-92-


[ Technical Paper ]

NVH Analysis Using Full Vehicle Multi Body Dynamic Model*-Influence of Constant Velocity Universal Joints on Shudder Vibration–

1. Introduction

In developing automobiles, the mode of NVHproblems 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 carrequire many hours of work and pose a significantchallenge in expediting vehicle design completion.Furthermore, since the NVH problems of a givenvehicle are often the result of complex mechanisms,determination of the root causes of each NVH problemthrough experiments with actual vehicles is verydifficult.

Recently, various structural and mechanismanalysis techniques have been applied to vehicledevelopment, 1) contributing to the simplification ofprototyping and reduction of labor time required forvehicle development. As for NVH problems, examplesof analysis techniques applied to idling vibration

analysis have been reported. 2) Analysis techniquesare used not only for vehicle development but also forautomotive component development. 3) 4)

We have developed a full vehicle dynamic modelusing mechanism analysis to analyze vehiclevibration. The NVH problem we analyzed this timewas the shudder vibration that occurs when a vehiclestarts and runs straight ahead. This full vehicledynamic model is a detailed mechanism analysismodel that incorporates virtually all the elementsusually included in automobiles, including a powertrain ranging from the engine to tires as well as asteering and suspension system. Using this fullvehicle model, we have reviewed the effects of a driveshaft constant velocity joint on shudder vibration. Theconstant velocity joint model also reflects nonlinearelements in its mechanism simulation. We haveverified that this model reproduces the forces that acton 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.

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

Yoshihiko HAYAMA****Takashi NOZAKI****

Masafumi NAKAKOUJI****Satoshi FUJIKAWA****

Komaki FUKUSHIMA****

NVH Analysis Using Full Vehicle Multi Body Dynamic Model

2. Shudder vibration in startingvehicles

Shudder vibration occurring in a starting vehicle isgenerally defined as the phenomenon in which a carbody is excited by resonance between the enginevibration and the induced thrust on a sliding constantvelocity joint.

Results of shudder vibration measurementsobtained 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 vehicleaccelerated while vibrating left and right. Theconditions that induced shudder on the test vehiclewere 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 driveshaft 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).

As shown in Fig. 2, the major component of theshudder 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 sideconstant velocity joint is responsible for shuddervibration on starting vehicles.

A non-mass-production sliding constant velocityjoint was used in our experiment in order to analyzethe shudder vibration phenomenon accurately.

Fig. 1 Experimental Results of Shudder Vibration

Fig. 2 Tracking result of fig.1 (c)

6050403020100

1500

500

0

1000

0 1 2 3 4 5Time sec

0 1 2 3 4 5Time sec

Spe

ed k

m/h

Torq

ue N

m

(a) Vehicle speed

(b) Drive shaft torque

1

-1

0

0 1 2 3 4 5Time sec

Acc

eler

atio

nm

/sec

2

(c) Acceleration of floor(Lateral direction; 50Hz LP Filter)

2000

1

250 300 350

Drive shaft speed rpm

Over allDrive shaft 3rd order

Acc

eler

atio

nm

/sec

2

-93-

3. Full vehicle model

3.1 Power train and chassis modelTo theoretically forecast the shudder vibration

phenomenon, analyses of engine vibrationcharacteristics and induced thrust on a constantvelocity joint, as well as transmission of the enginevibration and the induced thrust to the car body via themount and suspension, are also necessary. To thisend, we developed a detailed full vehicle mechanismsimulation model and performed analysis by applyinga drive torque from the engine output in order tosimulate an induced thrust occurring from rotation ofthe constant velocity joint.

Fig. 3 schematically illustrates the power train andchassis model used for our analysis work. The enginemodel simulated the piston and crank motion resultingfrom 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 suspensionresonance characteristics, and a suspension masscapable of simulating the rigidity of the stoppers on astarting vehicle. Furthermore, the tire model accuratelysimulates tire rigidity, thus allowing the simulation of acomplete vehicle running on a road, including frictionwith the road surface.


-94-

range of approximately 11.5 to 16.5 Hz. To investigatethe vehicle vibration mode in this frequency range, weperformed model analysis with the full vehicle model.

Fig. 6 illustrates the lateral direction engine mode.The characteristic frequency of the lateral enginemotion was approximately 15 Hz. The drive shaftspeed 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.

Fig. 3 Multi body Powertrain & Chassis model

Fig. 3 Acceleration of floor (Lateral direction)

200-1

1

1

0

0

250 300 350


Drive shaft 3rd orderExperimentSimulation

200 250 300 350


Acc

eler

atio

nm

/sec

2

Acc

eler

atio

nm

/sec

2

Fig. 4 Multi body CVJ model

(a) Inboard CVJ (TJ)

(b) Outboard CVJ (EBJ)

Our vehicle model was developed withconsideration for not only shudder vibration but alsopossible application for evaluation of other vehicleNVH performance. Thus, the vehicle model includesdetailed simulation of sections not associated withprediction of shudder vibration.

3.2 Drive shaft modelAs described in Sec. 2 above, we learned that the

drive shaft 3rd order component affects shuddervibration. The test vehicle has the TJ on its inboardside. In order to simulate the drive shaft 3rd order, theconstant velocity joint model must be a detailedmechanism 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)

4. Verification of analysis accuracy

Fig. 5 shows a comparison between the analysisresults 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.

5. Analysis of shudder vibrationphenomenon

5.1 Relation between engine and drive shaftWe found that shudder vibration on a starting

vehicle is related to the drive shaft rotation 3rd ordercomponent. The speed of a drive shaft that triggersshudder vibration is in the range of 230 to 330 rpm.The frequency of the 3rd order component falls in a

-95-

Fig. 8 Effect of engine mode on shudder vibration

200

1

0250 300 350


Acc

eler

atio

nm

/sec

2

Drive shaft 3rd orderOriginalModified

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

200

1

0250 300 350


Acc

eler

atio

nm

/sec

2

Drive shaft 3rd orderCVJ ACVJ B

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

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

Because of this, shudder vibration seemed to occurfrom the resonance between the lateral acceleration ofthe 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 lateralshudder analysis. Fig. 7 shows an engine vibrationmode with reduced engine mount rigidity. Thealteration to rigidity can be understood to change thecharacteristic vibration frequency and mode. Fig. 8offers a comparison of the magnitudes of shuddervibration between these two modes. In summary, foreach particular engine mode, there is a uniquevibration frequency and a unique magnitude ofshudder vibration.

5.2 Influence of CVJ typeInduced thrust, which is one of the drive shaft

rotation 3rd order components, typically occurs withthe TJ. When the TJ is run, an induced thrust occursalong the rotational axis of the outer race and thatmainly consists of the rotational 3rd order. Thisinduced thrust is a force that results from frictionalforces occurring because of the mutual friction of therotating internal components on the TJ.5) Therefore,we executed shudder vibration analysis with twosliding constant velocity joints with unique frictioncharacteristics. Fig. 9 summarizes the results ofanalysis for lateral acceleration below the floor at thedriver's seat. As can be seen from this diagram, itseems that use of a low friction CVJ can alleviateshudder vibration.


-96-


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 measurementvalues greatly varied. Fig. 10 illustrates the runningpattern used in the shudder test and the resultantmaximum measurements of shudder vibration. Themeasurement operation was repeated 30 times foreach constant velocity joint type. Each CVJ typeexhibited measurement variation and had uniquevibration characteristics. Notwithstanding, the lowestshudder measurements for both CVJ types werenearly same.

There are right-hand drive and left-hand driveshafts. The right hand and left hand TJs areincorporated via the differential box. The resultantforce of the induced thrusts from the two TJs variesdepending 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 gapangles between the two TJs. As can be seen in Fig.11(b), when the relative phase of one TJ coincideswith that of the other TJ, the resultant force of theinduced thrusts theoretically becomes zero, andreaches the maximum level when the phase gapangle is 60 degrees.

With our full vehicle model, the effect of the phasegap between the right-hand TJ and the left-hand TJwas analyzed. As shown in Fig. 12, it was foundthrough simulation that the magnitude of shuddervibration 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

Fig. 10 Mesurement summary of shudder test

1

0CVJ A CVJ B

Acc

eler

atio

nm

/sec

2

Drive shaft 3rd orderCVJ ACVJ B

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

(a) Running pattern

Takingoff

Takingoff

Stop

Stop

Turn Turn

Acceleration

AccelerationBraking

Braking

Fig. 12 Effect of CVJ phase gap on shudder vibration

00

1

30 60

Phase gap angle deg.

Acc

eler

atio

nm

/sec

2

90 120

ExperimentSimulation

Fig. 11 Theoretical effect of CVJ phase gapon induced thrust

Left Hand TJ

Right Hand TJ

(a) Definition of phase gap

(b) Induced thust

0 30 60

Phase gap angle deg.

TrunnionAxis (RH)

Dash Line:Right Hand TJ

Solid Line:Left Hand TJ

Rear Sideof Vehicle

Front Sideof Vehicle

Phase gap

TrunnionAxis (LH)

3rd

orde

r co

mpo

rnen

tof

indu

ced

thru

st

N

90 120

is apparent from Fig. 12, this trend was validated inthe 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 gapbetween the right hand and left hand TJs varied.

-97-

6. Conclusion

We developed a full vehicle running-capablemechanism simulation model and performed asimulation for vehicular NVH (analysis of shuddervibration).

We identified and verified the relationship betweenthe shudder vibration mechanism on a starting vehicleand engine vibration modes.

We qualitatively and quantitatively clarified theinterrelation between induced thrusts and shuddervibration levels on constant velocity joints.

References1) Kurisu, Fujikawa, Miyauchi, Koizumi, Hirobe,

Fukushima: Application of ADAMS mechanismanalysis software to power trains, Mazda TechnicalReview, No. 22, p44-49 (2004).

2) Matsumoto, Suzuki, Kitani, Kanuma: Idling shudderanalysis using a full vehicle model, the Society ofAutomotive Engineers of Japan, 2003 SpringScientific 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, p393-396 (2003).

4) Hayama: Internal force analysis for constant velocityjoints, the Society of Automotive Engineers of JapanJournal, Vol. 34, No. 4, p157-162 (2003).

5) Nozaki, Kobara: Experimental determination of forcesacting on tripod constant velocity joint housings, theJapan Society of Mechanical Engineers, 2003 annualmeeting proceedings, IV, p155-156 (2003).

Yoshihiko HAYAMA



Satoshi FUJIKAWA

Mazda Motor CoprorationPowertrain AdvanceDevelopment Dept.

Photos of authors

Takashi NOZAKI


Komaki FUKUSHIMA

Mazda Motor CoprorationPowertrain AdvanceDevelopment Dept.

Masafumi NAKAKOUJI



-98-

NTN TECHNICAL REVIEW No. 73（2005）

[ Technical Paper ]

Development of a High Precision Angle Sensor

1. Introduction

Rotary encoders for rotational angle detection havebeen used in a variety of industrial applications,including detection of actuation and travel amountsand motor speed control. In automotive applications,rotary encoders are utilized for steering system angledetection, accelerator pedals and throttle valves toimprove vehicle operability and safety. For everyautomotive application, the need has been mountingfor more compact rotational angle detection sensorsthat are capable of higher detection accuracy and areless expensive.

The detection schemes of rotary encoders can becategorized as contact and non-contact types.Components on a contact type rotary encoder candeteriorate from wear. Non-contact rotary encoderscan be further subcategorized into optical rotaryencoder types that utilize transmission or reflection oflight beams and magnetic rotary encoder types that

detect magnetic field variations. Since they are lesssusceptible to the effects of dust, magnetic rotaryencoders are used more often in automotiveapplications.

On many commonly used magnetic rotaryencoders, a magnetic sensor, such as a Hall sensor oran MR device, is used to detect the magnetic fieldvariation 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 smallfor the magnetic sensor to detect magnetic fieldvariations. 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 theShizuoka University Research Institute of Electronics,

***New Product Development R&D Center Mechatronics Research Dept.***Shizuoka University Research Institute of Electronics Professor

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 are

used 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 magneto-

resistive devices or Hall devices combined with a multi-pole magnet have difficulty in meeting these

requirements 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.

Toru TAKAHASHI****Yoshitaka NAGANO****

Shoji KAWAHITO****


NTN has developed a unique compact magnetic arrayrotary encoder that measures 5 mm square and iscapable of detecting an absolute angle of ±0.36˚. Atthe International Solid-State Circuits Conference 2005(ISSCC 2005) held in February 2005, this sensor wonthe Beatrice Winner Award for Editorial Excellence.1)

This paper provides an overview of our sensor andreports the results of angle detection with a compactbearing that incorporates this sensor.

2. Magnetic array sensor operatingprinciples

2.1 Structure of sensorThe rotary encoder we have developed consists of

a permanent magnet (using sintered NdFeB) at the tipof a rotary shaft and a 5 mm square sensor chipopposite the permanent magnet with a gap ofapproximately 0.5 to 1 mm between them (Fig. 1).

The internal configuration of the sensor circuitformed on the sensor chip is schematically illustratedin 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-and-hold amplifier (SHA) circuit that amplifies sensor

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Fig. 1 Rotary encoder setup

Fig. 2 Block diagram of the sensor chip Fig. 3 Structure of MAGFET (unit μm)

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

2.2 MAGFET magnetic sensor elementOur sensor chip uses a split-drain MAGFET

(Magnetic Field Effect Transistor) for each magneticsensor element. MAGFETs can be manufactured by aCMOS (Complementary Metal-Oxide Semiconductor)process similar to that used for ordinary logic circuits.They can also be incorporated into chips that alsoinclude amplifier and calculator circuits.

The MAGFET structure and corresponding circuitare illustrated in Fig. 3.

Like common transistors, each MAGFET has fourterminals, including a source (S), a gate (G) and twoseparate drain (D) terminals. This sensor is sensitiveto magnetic fields that are perpendicular to the sensorelements, and functions because of the Lorentz forceused 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=I2holds. When a magnetic field Bz is applied, electronswill flow in the direction of the red arrow in Fig. 3, andthere will be a bias between the currents of both drainterminals, resulting in a difference between the draincurrents (I1-I2) that occur in accordance with themagnetic field applied.

When assuming that the magnetic field intensity isBz [T], the sensor sensitivity is SR [1/T] and the sensorbias current holds a relation of lbias = I1+ I2, then thesensor output signal can be defined as follows:

Isignal = I1－I2

= SR BZ Ibias……………………………（1）


2.3 Angle detection schemeThe 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. Theblue line in this diagram shows the border betweenthe N-pole and the S-pole. The magnetic fieldperpendicular to the page from this line is zero, so thisline is called the "Zero-Line."

When the sensor signals from the multipleMAGFETs 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 ontheir coordinates.

θ= tan-1 (Δy/Δx)…………………………（2）

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Fig. 5 Output signal of the sensor arrays

Boundarybetween poles

Sensor signal

N-pole

S-pole

Two points where no magnetic field exists are detected.→ Boundary angle is calculated.

Fig. 6 Detection of the Zero-Line

Fig. 7 Block diagram of the angle calculator

To improve the detection accuracy of two zero-crossing points in the actual calculation process, theZero-Crossing Window, the area subjected tocalculation 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. Thesensor 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 calculatorcircuit, the positions of the two zero-crossing pointsare determined accurately to enable linearapproximation 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 10-bit angle data is output to an external system. Sincethis calculator circuit includes a communication circuit(Comm.I/F), it can output angle data through a serialcommunication means.

Fig. 4 Arrangement of sensor arrays(magnet is set at an angleθ)

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Fig. 8 Cross section of native-MAGFET

2.4 Features of our sensor systemOur sensor system calculates an angle based on

magnetic field distribution measured by multiplemagnetic sensors, and has the following advantages.(1) Since it is intended to detect a position where a

magnetic field does not exist (position where thelevel of magnetic sensor output is zero), our sensorsystem is not affected by sensitivity variation or thenon-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 isnot much affected by the positional differencebetween the magnet and the sensor elements.

(4) The MAGFET magnetic sensor chip can bemanufactured by a common CMOS process. As aresult, related circuits can be incorporated into thesame chip, reducing manufacturing costs.

3. Native-MAGFET

As explained above, our MAGFET magnetic sensorchip can be manufactured by an ordinary CMOSprocess. 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 areunavoidable because the chip is manufacturedthrough an ordinary CMOS process.

To attain sufficiently accurate angle detection withour sensor system, it was necessary to improve thesensor performance by enhancing the magneticsensitivity of the sensor elements. To achieve this, weinvented a native-MAGFET structured as shown inFig. 8.3) This diagram shows cross-sectional views ofCMOS 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 concentrationdirectly beneath the gate differs from that of a normal-

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

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

(1) Compared to the p-well region, the impurityconcentration 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 ionimplantation process, the concentration ofimplanted ions can vary, which leads to variation inthe characteristics of the sensor elements formedin this region. Native-MAGFETs are free from thisproblem, so variation of characteristics isminimized.

The results of measured magnetic sensitivities ofnormal and native MAGFETs are plotted in Fig. 9. Inthis diagram, IMAGFET represents bias currentscarried by the MAGFET. While the magnetic sensitivityof the normal-MAGFET was 2.7%/T, that of the native-MAGFET 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 native-MAGFET decreased as the bias current increased.Because the sensor elements were formed directly onthe p-substrate, the channel structure of regions withlower gate voltage seems to be similar to that ofembedded channels. We believe that with a highergate voltage, the distribution of channels shifts to thesurface. Since the channels are affected by diffusionat the interface, decreasing electron mobility,magnetic sensitivity is reduced.


4. Verification of our prototype chipoperation

The prototype sensor chip is shown in Photo 1.This chip was fabricated through a standard 0.25μmCMOS process [5-layer metal wiring, 1-layerpolysilicon, MiM (Meal-Insulator-Metal) capacitor]. Thesensor chip measures 5 mm by 5 mm, and the sensorarray measures approximately 4.2 mm square.

The technical data of the prototype sensor chip areas follows.(a) Number of sensor elements: 184 pixels×4 lines(b) Structure of one pixel: six MAGFETs connected in

parallel(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

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

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.

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Photo 1 Micrograph of a prototype chip

MAGFET Array

MAGFET Array

MA

GF

ET

Arr

ay

MA

GF

ET

Arr

ay

AngleCalculator

SHA-ADC

SHA-ADC

SH

A-A

DC

SH

A-A

DC

Photo 2 Picture of a sensor bearing sample

Fig. 11 Cross section of sensor bearings

(a) Conventionalsensor bearing

(b) Newly developedsensor bearin

Sensor chipHall sensor

Multi-pole ring Permanent magnet

Fig. 10 Angle detection result

Sensor

6. Conclusion

We have developed a novel rotational angle sensorthat employs magnetic sensor arrays. To improve thecharacteristics of sensor elements on siliconsubstrates, we developed a new sensor structureusing a native-substrate MAGFET to enhance themagnetic sensitivity of the sensor elements, therebyachieving a very high angle detection accuracy of 10bits per revolution (0.36 degrees). This degree ofangle detection accuracy was also achieved with asensor bearing that incorporates a MAGFET sensorchip sealed in a compact package.

Automotive components are becoming increasinglymore compact and functional. We intend to apply oursimple, compact sensor chip to automotiveapplications including steering systems andaccelerator pedals.

References1）S. Kawahito, T. Takahashi, Y. Nagano, and K.

Nakano, "CMOS Rotary Encoder System Based onMagnetic Pattern Analysis with a Resolution of 10bper Rotation," ISSCC2005, Dig. Tech. Papers, IEEEInt. Solid-State Circuits Conf., pp.240-241, 2005.

2）K. Nakano, T. Takahashi, S. Kawahito, "AngleDetection Methods for a CMOS Smart RotaryEncoder," Journal of Robotics and MechatronicsVol.17(4), pp.469-474, 2005.

3）Toru Takahashi, Kazuhiro Nakano, and ShojiKawahito, "CMOS Magnetic Sensor Arrays UsingNative-Substrate MAGFETs for a Smart RotaryEncoder," Proc. IEEE SENSORS 2004, pp.973-976,2004.

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Fig. 12 Angle detection result using a sensor bearing

Det

ectio

n er

ror

[deg

rees

]

Magnet rotational angle [degrees]

Toru TAKAHASHI



Photos of authors

Yoshitaka NAGANO



Shoji KAWAHITO

Shizuoka UniversityResearch Institute of Electronics

Professor


The optimal resolution with conventional sensorbearings was approximately 2 degrees (180pulses/revolution). As pulse output types, they cannotdetect absolute angles. In contrast, our sensor bearingis 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. Theangle detection accuracy of the new sensor bearing isplotted in Fig. 12. The detection error at a samplingspeed of 10 kHz was within ±0.3 degrees.

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[ New Product ]

Tapered Roller Hub Bearings for Large Truck

1. Introduction

Recently, unit designs have been increasinglyimplemented in front axles for large trucks in order toimprove reliability and assembly ease. Previously, twosingle-row tapered roller bearings had been used forthis purpose. However, this arrangement occasionallyposed safety problems. Incorrect adjustment ofpreload on an assembled bearing during truck repaircaused 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 unitdesign axles, and now some Japanese truckmanufacturers are using 2nd generation tapered rollerhub bearings (hereafter, GEN2 HUR). This paperdescribes the current trend in axle bearings on largetrucks as well as the GEN2 HUR that NTN hasrecently developed. This unit consists of an outer raceintegrated with a hub ring and has the largest flangediameter in the world.


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

Hiroshi KAWAMURA**Akira FUJIMURA**

2. Construction and features of largetruck axles

Previously, two single-row tapered roller bearingswere usually used for large truck axles. Then, GEN1units composed of two bearings and set light bearingswere adopted. Compared to previous designs that usetwo single-row tapered roller bearings, these designsboast higher reliability and ease-of-assembly. In 2000,NTN began mass-production of GEN2 HUR for thefirst time in Japan. This bearing type includes a unitcomprised of a bearing and a hub ring mountingflange (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 hasdeveloped a GEN2 HUR that features the world'slargest flange diameter. Furthermore, to attain higherreliability, NTN is currently developing GEN3 HURsthat do not need axle nut tightening control.


Table 1 Features of Tapered roller hub bearings in each generation

Table 2 GEN2 HUR Mass production results

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Generation

Rotation system

Assembly and service ease

Compactness

Built-in seal

Preload control

Generation

Rotation system

Assembly and service ease

Compactness

Built-in seal

Preload control

Ranking：☆＜☆☆＜☆☆☆

Driven wheel use

Non-driven wheel use

Conventional GEN1

Conventional GEN1 GEN2 GEN3

GEN2

Outer race rotation Outer race rotation

Outer race rotation Outer race rotation Outer race rotation Inner ring rotation

Outer race rotation

－ ☆

－ ☆ ☆☆ ☆☆☆

☆☆

－ ☆

－ ☆ ☆☆ ☆☆

☆☆

－ ☆☆

－ ☆☆ ☆☆ ☆☆

☆☆

－ ☆

－ ☆ ☆☆ ☆☆☆

☆☆

Front wheel (non-driven wheel)



Trailer rear wheel (non-driven wheel)

Truckmanufacturer Truck type Application

Beginning ofmass-production

Company A (Japan)

Company B (Europe)

Company C (Europe)

Large truck(24 t) 2000～

2001～

2002～

2003～

Large truck(24 t)

Medium truck(12 t)

Trailer

3.2 GEN2 HUR DevelopmentNTN has developed a large size GEN2 HUR that

can be used not only for large truck front axle non-driven wheels, but also for driven wheels on rear fullfloating axles. We will describe this newly developedproduct 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.

(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 raceintegrated with a hub ring, the newly developed HURhas been optimally designed to reduce the stress thatoccurs on the hub due to high load when the truckturns as well as to lighten the bearing. Fig. 3 showsan example of GEN2 HUR application. The assembly

3. NTN HUR development status

3.1 Mass-production of GEN2 HURGEN2 HUR technology is described in NTN

Technical Review No. 70. NTN's GEN2 HUR mass-production record is summarized in Table 2.


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

(2) Hub ring, inner race and roller materialsThe hub ring is made of a unique NTN bearing

material that features good castability and excellentshock resistance. Areas where greater stress canoccur when the bearing operates are treated with aninduction heating process to give them higher staticstrength and rotary bending fatigue strength. Thematerial for the inner ring and rollers is SUJ2.

(3) LubricantFor axle bearings on large trucks, we are adopting a

newly developed grease that boasts frettingresistance, 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 preventoverheating and scuffing-induced wear that tendsto occur on inner race larger ribs.

2 With a urea-based thickener featuring excellentshear stability, the grease boasts good durability.

HUR

Shaft

Knuckle

Nut

<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

1328.3

φ25

2

φ50

Hub ring

Roller

Inner raceCage

Fig. 1 Front non-driven GEN2 HUR structure

75.6559.5

φ60

φ25

2

Fig. 2 Rear driven GEN2 HUR structure

Table 3 General characteristics of greaseThickenerBase oilOperating temperature rangeHue

Urea-based thickenerMineral oil + synthetic oil

－30～＋150˚CYellow

Table 4 Test result of grease fatigue

Grease type

Current representative greasefor trucks (Sunlight 2)

HUR grease

Life test result (hr)

46

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

(4) CageThe polyamide resin cage used can be easily

assembled into the bearing and has excellent heatresistance, flexibility and durability. This makespossible a lighter weight at a lower price. Its shapehas also been optimized to promote ease of assemblyinto the bearing.

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(6) Hub ring (outer race), inner race racewaysurface, and roller rolling surface and largeend face forms

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 raceraceway surfaces, as well as the roller rolling surfaces,have optimal complex crowning (example of contactstress 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 resultAxle 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 rigsfor evaluating axle bearings, and can evaluate HURsfor large trucks through various tests such as rollingfatigue life test. One of the NTN test rigs and a part ofthe HUR evaluation results are described below.

(1) Evaluation test rigThe NTN durability test machine for large trucks is

schematically illustrated in Fig. 5. This machine iscapable of applying a maximum radial load of 196 kNand a maximum axial load of 98 kN. In executingtests, 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, itis capable of simulating cyclic loads or actual vehiclerunning modes.

Outer side Inner side

Evaluationthreshold

Con

tact

str

ess

Outer side inner raceOuter side outer raceInner side inner raceInner side outer race

Fig. 4 Example of contact stress calculation

Table 5 Test result of seal performance against mud water

2

4～7

18～20

30～38

Seal type Cross-section Mud water resistance (durability cycle)

Durabilityratio

Conventional seal

Double-lip seal

High-pack seal

New seal (double-lip high-pack seal)

1

2～3.5

9～10

15～19

※ 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

(5) SealA double-lip high-pack seal has been adopted to

provide good seal performance against mud water,which trucks traveling long distances oftenexperience, and to simplify the bearing assemblyprocess. Table 5 summarizes the results of a mudwater resistance test with various seal types. The sealmaterial for our newly developed bearing is fluorinerubber (FKM) to ensure better high temperaturedurability. The slinger material that provides the slidingsurface of the lip is stainless steel, making it rust-resistant.


weight reduction of approximately 20%.

1 Hub ring stress analysisPredetermined moment loads were applied to

various areas, a set number of times, to measure thestress occurring in each area. We verified that eacharea successfully withstood load application and didnot reach the fatigue limit. An example of hub ringstress analysis is provided in Fig. 6.

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Fig. 5 Test machine structure

RADIAL LOADHYDRAURIC CYLINDER

AXIAL LOADHYDRAURIC CYLINDER

ELEC. MOTOR

ACCELERATIONSENCER

TIRE RADIUS

TEST BEARINGTEST BEARINGTHERMO COUPLE

Fig. 6 FEM analysis example

(2) HUR analysis and testing results The results of analysis and testing of the HUR

illustrated in Fig. 2 are described for threeaspects–hub ring stress analysis, heavy load durabilityand 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 levelsbelow standard values, thereby we have succeeded in

(Before weight reduction) (After weight reduction)

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BearingNumber ofrevolutions

L（×104rev）

Calculatedlife

L0（×104rev）

Calculatedlife ratio

L/L0State after test

Frontnon-drivenwheel

Reardrivenwheel

2545

2545

2545

867

2382

867

3.0

3.0

3.0

3.0

8.2

3.0

Suspended

Suspended

Suspended

Suspended

Suspended

848

289

Flaking on innerside inner race

Table 6 Test results of 0.6G turning moment durabirity

Bearing Speed（r/min）

Runningtime（h）

State after testTest result

326(vehicle speed

160 km/h)

Suspended

Suspended

n=2No problems

such as seizure

n=2No problems

such as seizure

20

Frontnon-drivenwheel

Reardrivenwheel

Table 7 Test results of high speed durabirity

Photos of authors

Hiroshi KAWAMURA


Akira FUJIMURA


2 Heavy load durability test (0.6 G turningmoment durability test)

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.3 High speed durability test

The results of a high speed durability test aresummarized in Table 7. Even if the bearing wasoperated 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 hassufficient durability.

4. Conclusion

NTN has been mass-producing GEN2 HURs asaxle bearings for large trucks, and has recentlydeveloped a unique GEN2 HUR product that is thelargest in the world. Furthermore, NTN is currentlydeveloping GEN3 HUR products that boast greatlyenhanced reliability. We believe that these HURproducts are excellent axle bearings capable ofsatisfying requirements for long life, improvedreliability and ease-of-assembly. We expect thedemand for them will steadily increase in Japan. NTNis also mass-producing HURs for driven wheels forrear semi-floating axles on compact commercialvehicles. We intend to supply various HUR products tosatisfy user needs.


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The Auto-tensioner Market and Technical Trends

1. Introduction

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


Automotive engines generally use accessory drive belts, timing belts, or timing chains to transmitcrankshaft 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 asreducing noise generation.

Although the cost of the system and the number of components increases when auto-tensionersare 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 toimprove 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 reducecost and increase reliability.

Satoshi KITANO**Tadahisa TANAKA**

Tomokazu NAKAGAWA**

2. Drive systems

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.


3. History of camshaft drive systems

The first timing belt drive system in the world wasused commercially in 1946 in the USA, and the firstautomotive 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 systemwas the main type, but in the 1980's, timing beltsystems became dominant. However, timing chaindrive systems began to be adopted again in the1990's thanks to the development of high-performancechains. The current market shares for timing chaindrive systems and timing belt drive systems are about50% each in Japan.

Fig.2 illustrates the direction of camshaft drivesystems.

In the past, all accessory drive systems involvedseveral belts. In the 1980's, though, serpentinesystems began to be used in the USA. The market

Fig. 1 Layout of Auto-tensioner (example)

Fig. 2 Direction of a cam-shaft drive system

-111-

20000

10

20

30

40

50

60

70

100

80

90

2001 2002 2003 2004 2005 2006

Year

Per

cent

age

of v

ehic

les

Others

Timing chain drive systems

Timing belt drive systems

Auxiliary drive belt use (serpentine)Timing chain useTiming belt use

Camshaft Camshaft

Idler

Idler

Idler

Air-conditioner

Tensioningpulley

ActuatorCrank Crank

Generator Water pump

Auto-tensioner

Crank

Chain guide

Chain guide

Direction ofengine shaftrotation

Direction ofengine shaftrotation

Cam sprocket

Chain tensioner

Supportingpoint

Direction of engine shaft rotation

Auto-tensioner

share for this drive system type is expanding in Japancurrently.

4. Importance of auto-tensioners

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


-112-

Fig. 4 illustrates various timing belt auto-tensioners.NTN has been developing various hydraulic

dampers including a standard type, a short type that iscompact in the axial direction, and an ultra-compacttype.

The integrated pulley-arm type is a unit designcomplete with a hydraulic damper and a tension pulleythat reduces costs and has smaller spacerequirements. The built-in pulley type helps savespace since the hydraulic damper is built into thepulley, and it can replace a fixed tensioner.

A performance comparison of various timing beltauto-tensioners is shown in Table 1.

5. Auto-tensioner types and structures

5.1 Timing belt auto-tensioners The actuation systems for timing belt auto-

tensioners include hydraulic, screw, and frictionsystems. NTN auto-tensioners employ a hydraulicactuation system that excels in durability and tensionadjustment capability to cope with temperaturevariations.

Fig. 3 illustrates the structure of an NTN auto-tensioner. The cylinder is die-cast aluminum to reduceweight.

Miniaturization Miniaturization

MiniaturizationIntegration

Standard type

Separate type Integrated pulley-arm type Built-in pulley type

Ultra-compact typeShort type

Unit design

Tension spring

Hydraulic damper

Fig. 4 Various Auto-tensioner's for timing belt

Fig. 3 Sectional view of Auto-tensioner

Table 1 Performance comparison (for timing belt)

Cylinder (die-cast aluminum)

Hydraulic fluid

Top level of oil

Air

Wear ring

Plunger

Valve sleeve

Valve retainer

Steel ball

Separator

Plunger spring

Return spring

Rod

Oil seal

Retainer ring

Pin

Hydraulic type

Hydraulic type

Screw type

Viscosity type

Friction type

Type Manufacturer Function※

※Tension adjustment function

Reliability Weight Cost

NTN

Ａ

B

C

NTN

D

E

F

◎ ◎ ◎ ◎

◎ ○ ○ △

○ ○ ○ △

△ △ ◎ ◎

◎ ◎ ○ ○

△ ○ ○ △

△ △ △ ○

△ △ ◎ ◎

Ele

men

tU

nit

-113-

NTN auto-tensioners employ a hydropneumaticconstruction that is free from oil leakage. However, ifthis type of auto-tensioner is mounted at an angleclose to horizontal, the liquid can approach thepressure chamber, possibly letting air enter andleading to damper failure. Therefore, the mountingangle of this auto-tensioner is limited, and therecommended mounting angle range is ±45˚ fromvertical.

If a greater mounting angle is unavoidable, an auto-tensioner 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.)

5.2 Timing chain auto-tensionersRecent camshaft drive systems employ timing

chains rather than timing belts so that the wholeengine length can be shortened and becausereliability is high and running noise is low with timingchains.

With ordinary timing chain auto-tensioners(hereafter referred to as chain tensioners), engine oilsupplied from the engine is introduced into the chaintensioner to generate hydraulic damping power. A no-back mechanism is included to prevent inwardpressure before engine oil is supplied after the engineis started. No-back mechanisms can be categorizedas ratchet, serrated screw, ring, and screw types.

NTN uses serrated screw and ring types. To reduceweight, die-cast aluminum cylinders are used. Massproduction 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. 7 FEM analysis result

PlungerPlungerScrew rod

Check valve

Relief valve

Relief valve

Ring Fixture

Rod seat

Assist spring

Return spring

Spring seat

Return spring

Serratedscrew Air expulsion

orifice

Air expulsion screw



Check valve

Fig. 8 Various Auto-tensioner for timing chain

Reservoir

Pressure chamber

Top level of oil

Auto-tensionermounting angle

Fig. 5 Auto-tensioner with reservoir room

Fig. 6 3D-model


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

2 Static no-back characteristicsWhen the engine is at a standstill, and if the

span where the chain tensioner is present is in atensioned state because of a phase differencebetween the camshaft and the valve, then thetensioner will be pressed in. Then, after the engineis started, and before the engine oil is supplied, theamplitude of vibration could be greater due to theamount of inward pressure, possibly causingabnormal noise. Fig. 10 summarizes measuredstatic characteristics of the no-back mechanism.

The results in this chart were obtained byallowing a static load to be applied to the plungertip and measuring the resultant plungerdisplacement until a particular load was reached.Though the amount of inward pressure relative tothe screw axial clearance is apparent, no furtherinward pressure occurred because the screwremained locked.

The results of a performance comparison of varioustiming chain auto-tensioners are summarized in Table2.

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 no-back mechanism can function regardless of the screwposition, actuation is stage-less, and the returnallowance 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 operateswiftly in the outward direction.

The results of an investigation into the functions ofserrated screws are described below.1 Results of plunger tracking verification

Any auto-tensioner must be able to quicklyabsorb changes in tension due to engine thermalexpansion or due to chain or belt slackening whenthe engine starts.

-114-


Table 2 Performance comparison (for timing chain)

Type Manu-facturer

Basic structureDamper No-back

Cost

NTN◎ ◎ △

○ ○ ◎

○ ○ ○

○ ○ ○

○ ○ ○

△ △ △

Hydraulic andserrated screw

Hydraulic

Hydraulic

Hydraulic

Hydraulic

Ring

Ratchet

Ring

Ratchet

Screw Screw

Serrated screw(stage-less)

A

B

C

D

Serratedscrew type

Ring type

Ratchet type

Ring type

Ratchet type

Screw type

Fig. 9 Plunger position during engine operation

Time elapsed min

Gui

de d

ispl

acem

ent

mm

Out

war

d di

rect

ion

6

5

4

3

20 5 10 15 20 25 30 35 185 190

Standstill (25°C)

3000r/min（25˚C）

3000r/min（80˚C）

500r/min（25˚C）

500r/min（25˚C）

500r/min（80˚C）

Standstill

Start-upRestart

Fig. 10 Static characteristic of 'No-back' mechanism

Screw axial clearanceMeasured three times

0.200.200.20

Plunger position change

Load（N）

3 Dynamic inward pressure characteristicsWhile the engine is running, the chain tensioner

plunger pushes in to prevent chain snaking, butover-tensioning the chain must also be avoided.

The results of measurements of the inwardpressing on the actuated chain tensioner aresummarized in Fig. 11.

-115-

As shown in this chart, when the chain tensioneris actuated, the screw is unlocked, and the plungeris pushed in, preventing the chain from becomingover-tensioned.

0

10

20

30

40

10.00

9.75

9.50

8 sec

80 loading cycles

Amount of inward pressing after 80 loading cycles：0.150mm

Frequency: 10 Hz

Dis

plac

emen

t m

mLo

ad

kgf

Fig. 11 Dynamic characteristic of screw plunger

2000

1500

1000

500

0

-5000 100 200 300 400 500

Gen

erat

ed lo

ad

N

Vibration frequency Hz

Maximum level

Minimum level

Plunger amplitude: ±0.05mm

Fig. 12 Dynamic-characteristic evaluation result

5.2.2 Damper adjustment mechanismThe chain tensioner receives a varying load from

the 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 forcesettings.

2 About trackingThe varying load acting on the chain tensioner is

governed by the number of engine cylinders.

The currently available NTN chain tensioners are forfour-cylinder and six-cylinder engines. With a four-cylinder engine, a rotational 2nd-order varying load isapplied to the chain tensioner, and with a six-cylinderengine, a rotational 3rd-order varying load is applied.With a six-cylinder engine, the maximum vibrationfrequency of a varying load can reach 300 Hz. Whenthe speed of a four-cylinder engine on a motorcyclereaches 15,000 r/min, the frequency of the varyingload can reach 500 Hz. Every chain tensioner must becapable 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 stablyto a maximum frequency of 500 Hz, and that the chaintensioner reliably responds to the varying load.


5.3 Auto-tensioners for accessory drive beltsPreviously, commonly employed belt tensioners

used multiple belts with one or two auxiliarymechanisms 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 onebelt because of reduced maintenance needs, ease ofbelt installation, reduced weight and shorter totalengine length. At first, serpentine designs were usedfor large gasoline engines with six or more cylinders,but they have been increasingly used for gasolineengines and diesel engines with four or fewercylinders.

However, the longer belts of serpentine systems areassociated with elongation that occurs when drivingthe 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 tensionchanges due to variations in the load from auxiliarymechanisms and in the engine speed, the auto-tensioner 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 greaterdamping force and higher reliability.

-116-


The construction of an NTN auto-tensioner isillustrated in Fig. 13.

A performance comparison of various auto-tensioners 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-relatedcomponents and performed FEM analysis to achieveboth reduced weight and higher rigidity.

In order to help promote environmentalconservation by improving automobile fuel economy,an idling-stop mechanism for automotive engines isavailable. With this scheme, the integrated startergenerator (ISG) drives the accessory belt to turn thecrank to restart the engine. NTN has been developinghydraulic type accessory belt auto-tensionersoptimized for ISG engines.

We believe that the demand for hydraulic auto-tensioners will be increasing because they will enablelonger drive belt life as loads from auxiliarymechanisms, including larger alternators and greaterair-conditioner capacities, increase.

LinkageCylinder (die-cast aluminum)

AirTop level of oilHydraulic fluid

Wear ring

Plunger

Valve sleeve (steel)

Linkage

Retainer ring

Oil seal

Rod

Return spring

Plunger spring

Check ball

Valve retainer

Spring seal (die-cast aluminum)

Fig. 13 Auto-tensioner for accessory belt

Table 3 Performance comparison (for accessory belt)

Hydraulic

Frictional

Type Manufacturer Function Reliability Weight Cost

NTN

Ａ

B

◎ ◎ ○ ○

○ ○ ○ ○

△ △ ◎ ◎

-117-

Photos of authors

Satoshi KITANO


Tadahisa TANAKA


Tomokazu NAKAGAWA


6. Conclusion

We have introduced the trend for various auto-tensioners related to improved reliability in drivingcamshafts and accessory shafts. We are determinedto develop less expensive, highly reliable auto-tensioners that feature improved functionality andconstruction.

References1）Kazuki Kawashima, NTN Technical Review No. 61

(1992)2）Katsumi Furutani, Kazuki Kawashima, NTN Technical

review No. 65 (1996)3）Satoshi Kitano, Tadahisa Tanaka, Tomokazu

Nakagawa, Monthly Tribology (2004, 10)


-118-


[ New Products Information ]

Advanced Technology Deep-Groove Ball Bearings forHigh-Speed Servomotors

Features

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

Higher speeds are capable with these prelubricateddeep-groove ball bearings!

0

2000

3000

1000

6000

5000

4000

15000 dn=600,000 dmn=950,000

120009000600030000

20

40

60

80

100

［］

Test

dur

atio

n h

Grease

Bearingdescription

Bearingdescription

6209 with non-contact seals on both sides（φ45mm×φ85mm×19mm）

RotationspeedBearingtemperature

Blue：n=11,111min-1，Red：n=13,333min-1

130˚C (bearing outer ring)

6308 with non-contact seals on both sides（φ40mm×φ90mm×23mm）

Rotation speed min-1Out

er r

ace

tem

pera

ture

incr

ease

˚C Rapid temperature increase test High-speed, high-temperature durability test

Steel plate wave-type cage+

MP-1 grease

New cage+

ME-1 grease

New cageCrown-shaped cage

ME-1 grease common to both cagespecifications

Newly-shaped synthetic resin cageHigh-speed, long-life grease

Capable of dmn value 95×104

PrelubricatedDeep Groove Ball Bearings

62XX and 63XX series

¡Optimal design improves lubrication reliability

New Products Information

-119-

We have developed a constant velocity jointthat features the lowest level of idling shudder in the world.

Idling Shudder-Suppressing FixedConstant Velocity Joint

Features

Performance

Application on actual vehicle

¡Compared to similar competitor products, thenew NTN fixed type constant velocity jointfeatures a 40% reduction in bending load inthe operating angle direction, and a 20%reduction in actual vehicle idling shudder.

20% reduction

EBJ-M(new CVJ)

Competitorproduct

Idling shudder (engine speed: 620 rpm)

Shu

dder

acc

eler

atio

n

Improvement by adoption of low slide-resistance CVJ (PTJ)

By means of bending rigidity-reduced CVJ

Differential gearing

Sliding jointFixed joint

Shaft

Absorption by sliding

Vibration from engine

Absorptionby bending

-120-

NTN has developed the world's thinnest low profilethrust needle roller bearing!

Low Profile Thrust Needle Roller Bearings

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

NW type (low profile cage)Roller diameter: between 1.5 mm and 2.0 mm

FW type (ultra-low profile cage)Roller diameter: between 1.0 mm and 1.5 mm

Press-flattening process for roller fixing parts

Single-sheethigh-strength cage

Roller guide section

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

W type (conventional cage)(New)

NW type (low profile cage)(New)

FW type (low profile cage)

●No interference between the roller fixing parts on the cage and the rollers

●Interference avoided by press- flattening 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

¡Reduction of moment load at stress concentration areas by low profile design

¡Reduction of moment load at stress concentration areas by increased sheet thickness and low profile design

●Interference is avoided by ironing the cage pockets

●The single sheet form has a low profile, as well as higher strength from increased 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

(Cage strength relative to W type)3.0

-121-

Sets new standards with longest life and greatest load capacity in the world!

S-TITAN™ Series of Self-Aligning Roller Bearings

Features

Ordinary temperature life

High temperature life

Load capacity comparison (dynamic loadrating) with competitor product

¡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.

500

400

300

200

100

Current Company A S-TITANTM0

Life

(h)

Test conditionsTest bearing:

22208(dia. 40×dia. 80×23)

Test load: Dynamic loadrating ratio = 0.5

Rotation speed: 2000 min-1

Lubrication: Turbine oil (oil bath)

3000

2500

2000

1500

1000

500

Current S-TITANTM0

Test conditions

No.

of l

oad

(cy

cle)

Test temperature: 200˚C

Test piece: dia. 47×7 mmflat plate

Contact stress: 5.5 GPa

Load frequency: 3000 r/min

Lubrication: Ether oil (oil bath)

Bore, mm0 50 150 250

(%) 160

150

140

130

120

110

100

90

Com

paris

on o

f dyn

amic

load

rat

ings

(N

TN

/Com

pany

A)


-122-


Bearing with Magnetic Field Resistant Rotary Sensor

Structure

Magnetic field resistance: three times that of previous products

Magnetic sensor

Magnetic encoder

Performance

0

1

2

3

4Resistance to external magnetic fields

Product on the market NTN product

Rat

io

Supply voltage

Allowable temperature range

Output types

Applicable bearing numbers

5～24V, DC

－40˚C～120˚C

Rectangular wave, open-collector

Phases A and B

6202～6209

Specifications

Features

¡Sensor can be built into a high-torque motor,allowing very compact systems

¡Sensor rarely malfunctions even inenvironments with strong external magneticfields

¡Built-in sensor allows greater compactnessand easier assembly

¡Sensor can detect rotation even at speedsnear 0 rpm

¡Sensor output signal can be mathematicallyprocessed, allowing calculation of rotationalangle and velocity and detection of rotationaldirection

-123-

First in the bearing industry! Development of an ultra compact size of dia. 14××12.Great space savings in office automation equipment.

"Ultra" Compact Torque Limiterfor Office Automation Equipment

Performance

3002502001501005000

5

10

15

20

25

30

35

40

Time h

Torq

ue

mN・

m

Conventional productNew product

<Test conditions>

Temperature: Ambient / room temperature

Operation pattern: Actuation for 1.1 seconds and no motion for 0.2 seconds

Rotation speed: 220 min-1

Torque stability is equivalent tothat of conventional products

Rubber roller diameter can be made smaller and the axialdirection more compact (red line shows a conventionaltorque limiter).

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).

Rubberroller

Rubberroller

Newproduct

Newproduct

Conventionalproduct

Conventionalproduct

Conventionalproduct New product

Dia. 14×12 mm (conventional product: dia. 18×18 mm)

Easy paper-jam maintenance

Ultra compact size

Bi-directional rotation capability


-124-


RustGuard™ Highly Corrosion-Resistant Bearing forSteelmaking Machinery

Application

Results of failure modereproduction test on actual machines

Special coating improves bearing corrosion-resistance and extends life to3.5 times that of conventional products.

Features

*Roller mark rusting: Each roll neck bearing on a rollingmachine is left stationary water present on it during rollermaintenance work. As a result, streak-patterned axial rustingoccurs at equal intervals (roller-to-roller pitches) on theraceway surfaces on inner and outer races. This phenomenonis called "roller mark rusting."

¡Use in steelmaking and papermakingmachinery parts that are exposed to water

Flaking

Roller markrusting

Evidence ofbearing failureon actualmachinery

<Conventional product vs. RustGuard>

1

5

10

106 107 108 109

20

50

80

90

Acc

umul

ated

failu

re p

roba

bilit

y %

Life (loading cycles)

RustGuardL10 =7.4×107 cycles

Conventional specification

L10 =2.1×107 cycles

¡Longer bearing life with improved corrosionresistanceWith 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 thickerthan that attained with other bearing products. In situationswhere corrosion (such as roller mark rusting*) can occur,RustGuard bearings boast a life 3.5 times longer thanconventional bearings.

-125-

Hybrid BEARPHITE Bearings

Structure

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

Sintered material:Copper-based

Sintered metal layer

Resin

Resin layer

φ16φ8

5Performance (frictional characteristicsfor 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)

0.3

0.25

0.2

0.15

0.1

0.05

00 100 200 300 400 500

Test time h

Fric

tiona

l coe

ffici

ent

μ

Hybrid BEARPHITEResinSintered bearingRolling bearing

Wear on sintered bearing 10μm

Wear on resin 19μm

Wear on hybrid BEARPHITE 6μm

Borediameter

Bearing outsidediameter Length Roller bearing number

1

2

3

4

φ 6

φ 8

φ 8

φ10

φ12

φ12

φ16

φ15

4

3.5

5

4

WBC6-12ZZA

W678ZZA

W688ZZ

W6700LLFV4

Dimensions

Features

¡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


-126-


Mono-Drive Two-Way Feeder™

Plan view

Breakthrough two-way feeding unit achieves parts storage, aligning andfeeding functions in one linear feeder!

Bowl system (conventional)

Mono-drivetwo-way feeder

Applications

Features

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

This novel return feeder is comprised of a newly developedleaf 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 lightweightCompared with conventional bowl feeders, this linear aligningfeeder 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 overallequipment weight reduction.

¡Simple construction makes it the optimumchoice for handling multiple work piece typesand small lot production

The simple construction allows easier maintenance. Thefeeder can handle multiple work piece types by simplyinstalling 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.

-127-

Ultra Compact Grinding Sludge Solidification System

Capacity and applications by type

Low power consumption, easy operation, and environmental impact alleviation!

Type

Ultracompact

Compact

Handling capacity for liquefiedgrinding sludge

10 kg/h (1.8 ton/month)※

20 kg/h (3.6 ton/month)※

Local processing (easy directconnection to grinding sludgefiltration systems)

※These values were obtained by assuming that the equipment is operated　8 hours per day, 22 days per month.

Centralized processing,batch processing

Applications

Dimensions and weight of ultracompact type

Features of ultra compact 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 deliveredsludge as small as 10 kg/h

¡Very economical at 50% the cost of, and usingless power than, the compact type

Both types can be used according to needs.

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

¡Cost reduction by recycling and reuse of oil-based coolantGrinding liquid recovery ratio of 90% orhigher

¡Reduction of industrial waste emissionsCapable of reducing sludge volume to 20%and weight to 55%

¡Reduction of waste transportation and landfilldisposal expenses

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


special issue; automotive products...ntn technical review no.73 ¢2005 £ < new product > high...

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