Special Issue
Special Supplement to Industrial Machines
April 2004

ISSN 0915-0528 OSAKA,JAPANCONTENTS

PrefaceTechnical
PapersWasaburo SUGANUMA

Improving Rolling Contact Fatigue Life of Bearing Steels Through Grain Refinement

12Chikara OOKI, Kikuo MAEDA and Hirokazu NAKASHIMA

Dynamic analysis of Cage behavior in a Cylindrical Roller Bearing

8

Tomoya SAKAGUCHI and Kaoru UENO

New
Products

Precision Bearings "ULTAGE" Series for Machine Tools

18

Futoshi KOSUGIBearings for High Speed CT Scanner

28

Yosuke OYA

Integrated Sensor Bearing Unit for Axleboxes

34

Masanori UENO

Bearings for Wind Turbine

40

Souichi YAGI

Insulated bearing "MEGAOHM" series
Hideji ITO

New High-Capacity HWTJ Type Pressed Cage and Needle Roller Assemblies
Katsufumi ABE

HK-F type drawn cup needle roller bearings

48

52

56

Hideki AKAMATSU

Hydrodynamic BEARPHITE Unit for HDD
Kiyotaka KUSUNOKIHigh Angle Active Link
Keisuke SONE, Hiroshi ISOBE and Koji YAMADATechnical
ArticlesImprovement of Leakage Magnetic Flux Resistance of Integrated Sensor Bearings
Takashi KOIKE, Tomomi ISHIKAWA, Hiroyoshi ITO and Noriyoshi MIZUTANIIntroduction of Grinding Swarf Recycling
Kanji NAKAMURA

New Products Information

62

70

74

80

84

NTN TECHNICAL REVIEW No.71（2004）[ Preface ]Special Issue

Special Supplement to Industrial Machines

Wasaburo SUGANUMA
Executive Director

In the second half of the 20th century, when industry was booming, awareness began toincrease about the effect that industrialization may have on the natural environment and on oursocial surroundings. It became clear that people strongly desired an environment where they
could live comfortably and safely. Now that we are into the 21st century, industrial products aredesigned with safety and human convenience as primary considerations.
As a result of these changes in the atmosphere, industrial machine designs are also
advancing to be more ecologically sound and people-friendly. For example, emphasis in themachine tool industry has been placed on considerations affecting the working environment--such as sound generation and air quality-- in addition to strictly performance-related issues
such as increasing speed and precision capabilities. The increased reliability of railroad
vehicles has been achieved simultaneously with increased speed performance andmaintenance intervals. Also, the use of sensors used for failure prediction, integrally designedwith the bearings used, is on the rise. CT scanners, imaging equipment used in the medical
field, are being widely used. To reduce patient stress during the exam, higher speed operationand lower noise levels are required by the scanner manufacturers. Furthermore, wind power
generation systems are being installed worldwide as a clean source of energy.
NTN has been advancing bearing designs in step with technological innovations in allindustrial machinery fields, and actively seeks improvement of various technologies for longer
life and reduction of size, weight, and friction for these applications. In this special issue, we
introduce technology for bearing life improvement, high speed and precision, andenvironmental preservation, together with the basic technologies that allow thesedevelopments. Included are the following subjects:¡FA-treated bearings, with drastically prolonged rolling contact fatigue life due to grain
refinement;¡ULTAGE series, which has gained a reputation for precision bearings in machine tools;¡Hydrodynamic BEARPHITE unit that is rapidly gaining acceptance in hard disk drive market;¡MEGAOHM series that features high insulation capability and reliability through theapplication of special ceramics and improvements in spray-coating methods
In the midst of global industrial structural changes, we hope that NTN can contribute, throughbearings and other precision products, to a global environment and living conditions in which
our children, who will be leaders in the coming decades, can live comfortably and safely.

-1-NTN TECHNICAL REVIEW No.71（2004）[ Technical Papers ]Improving Rolling Contact Fatigue Life of Bearing Steels
Through Grain Refinement

Chikara OOKI*
Kikuo MAEDA*
Hirokazu NAKASHIMA*

It is well known that yield strength can be improved by reducing the ferritic grain size (i.e. Hall-Petch’s Law) and thatthrough a reduction in the ferritic grain size an increase in a steel’s fatigue life is observed. However, the specific effectof grain refinement on rolling contact fatigue life has not been thoroughly investigated. The primary obstacle inresearching hardened steels (which possess a martensitic structure) has been the difficulty in obtaining small, uniform
grain sizes. Recently, some new methods for the grain refinement of hardened steels have been discovered. One suchmethod is called “Ausforming”, which induces a large deformation at temperatures over Ac1. Unfortunately, thisremarkable method is as yet unavailable for practical manufacturing. Keeping this in mind, we set out to develop a
specialized heat-treatment process. The main objective was to obtain a grain refined martensitic structure within JISSUJ2 (SAE52100 equivalent) bearing steel, produced by a standard manufacturing process, while at the same timeminimizing productivity loss. Thus far, we have succeeded in creating a prior austenite grain size of approximately 5 μm
in diameter (half that of the conventional grain size). It is believed that the packet or block size of the martensite
decreases proportionally with that of the prior austenite grain size. As a result, during RCF testing, the grain refined
SUJ2 material demonstrated a fatigue life that was twice as long as carbonitrized steel under both clean and debriscontaminated conditions. In addition, the grain refined SUJ2 material was superior in fracture strength and aged
dimensional stability. Therefore, it can be concluded that the grain refinement of existing bearing steels could prove tobe very useful in greatly extending fatigue life.

By and large, there have been few detailed studies
concerning the effects of grain refinement on steels'rolling contact fatigue (i.e. RCF) life. Primarily,
because it was believed that a) the grains ofquenched bearing steels were already considerably
small, b) excessively small grains hindered thehardenability of steels, and c) specialized heat
treatment processes such as "Ausforming" and the"Grange Method" 2) were required for further grain
refinement. This paper summarizes the attempt toextend the RCF life of bearing steels through grain
refinement via. a specialized heat-treatment process.

1. Introduction
Steel is a polycrystalline substance, containing
various microstructures such as prior austenite,martensite and ferritic grain boundaries. Many studies
have been conducted detailing the effects of grain size
variations on the macro-properties of steel. It is wellknown that yield strength of steel can be improved (aswell as, the fatigue life) through a general reduction in
the materials' grain size (i.e. Hall-Petch's Law) 1). This
technique is commonly referred to as "grain
refinement".

**Reserch & Development Center Technical Reserch Dept.

-2-

Improving Rolling Contact Fatigue Life of Bearing Steels Through Grain Refinement

after tempering at 180˚C for 2 hours). The steel
processed as a result of this specialized heat
treatment is hereafter referred to (within this study) as"FA-treated" material. Photo 1 shows the prioraustenite grain boundaries of a normally quenchedmaterial, carbonitrided material and FA-treatedmaterial. The FA-treated material possesses
crystalline-grain diameters that are at least 50%
smaller than that of the other two materials.

2. Refinement of Crystal Grains
Given that the size of crystalline grains are largely
dependant upon the quenching temperature 3), it isassumed that a more refined grain structure can beproduced if the quenching temperature is kept as low
as possible but still above the austenite transformation
temperature. However, to provide the steel with
sufficient hardness, it is necessary for at least aminimum amount of carbon to be dissolved inmartensite. Thus, the quenching temperature must be
determined by taking into account both the amount ofdissolved carbon and the degree of grain refinement
required. This is the reason why it is notconventionally possible to perform quenching at
considerably lower temperatures.
In this study, test specimens were initially nitrided toreduce the steel's Ac1 transformation temperature andmaintain a sufficient amount of dissolved carbon in themartensite during the primary quenching process. The
specimens were then heat treated at lower
temperatures (in the secondary quenching process) to
produce refined or smaller crystalline grains.
The adaptation of the grain-refining heat treatment
method allowed the JIS-SUJ2 (SAE 52100
EQUIVALENT) test material to form prior austenite
grain boundaries with an average size of 5mm or
smaller and hardnesses of HV700 or higher (even

3. Point Contact Type Rolling Contact
Fatigue Test
To investigate basic rolling contact fatigue
characteristics, cylindrical specimens (φ12×L22)
were tested under clean lubrication conditions. SUJ2
(SAE 52100 EQUIVALENT) material was used tofabricate these test specimens; one exhibiting anormal level of cleanliness, the other possessing a
lesser degree of cleanliness. Fig. 1 shows theanalytical results for the evaluation of non-metallic
inclusions, conducted using NTN's inclusion
quantification unit 4). Type! refers to A-type
inclusions with TiN, and Type@refers to B and C-type
inclusions without TiN. Regarding Type!and@, it was
shown that the specimen with the lesser degree of
cleanliness contained the greater percentage ofinclusions.

0.05mm

0.05mm10: type !
: type!
!

1

10-1

10-2

10-3

0

5

10

0.05mmCarbonitrided
Photo 1 The prior austenite grain boundaries
Number of inclusions（/mm2）

Number of inclusions（/mm2）

Normally quenched

15

20

Size of inclusions (μm) Cleanliness: Normal

FA treated10: type !
: type!
!

1

10-1

10-2

10-3

0

5

10

15

20

25

Size of inclusions (μm) Cleanliness: Less than Normal

Fig. 1 Size distributions of non-metallic inclusions (Measurement area:1000mm2)-3-

NTN TECHNICAL REVIEW No.71（2004）

Fig. 2 illustrates a schematic of the test rig, andTable 1 shows the test conditions.

4. Life Test Under ContaminatedLubrication Conditions
Since the cleanliness of steels has greatly improved
due to the advancement of steel manufacturing
processes, peeling that is due primarily to rolling
contact fatigue rarely occurs 5).
However, given the increasingly hostileenvironments in which bearings are being subjected(especially in automotive applications), the entry of
hard-particle debris is avoidable. It is clear that themain cause of damage to bearings today is due toflaking that starts from indentations formed by theseparticles. This section evaluates the benefits of FAtreated material within such applications.
L10 life testing was conducted with 6026-type ballbearings and 30206-type tapered bearings using
lubrication contaminated with relatively large hardparticle debris.

Driving rollSpecimen
φ12 × 22

Guide roll

3/4" ball
Guide rollFig. 2 φ12 point contact type rolling contact fatigue test rig
Table 1 Test condition of point contact fatigue test
Cylindrical specimen（φ12 x L22）
Specimen
3/4"（19.05mm）
Steel ballMaximum contact stress（GPa） 5.88
Load speed
（cpm）
46240
Lubricating oil
Turbine VG68, forced oil lubrication

4. 1 RCF-Test on 6206-Type Ball Bearing Under
Contaminated Lubrication Conditions
The bearings used for the test were normallyquenched, carbonitrided, and FA-treated 6206-typeball bearings manufactured from SUJ2 (SAE 52100
EQUIVALENT). Table 3 shows the metallurgicalconstituents from each bearing sample. The amountof retained austenite in the FA-treated bearing was
between those of the carbonitrided and FA-treatedbearings.

Table 2 shows the results of a point contact typerolling contact fatigue test. Regarding the steel
specimen possessing normal cleanliness, the L10 lifeof the carbonitrided material was found to be 3.1 times
longer than the normally quenched material, while theFA-treated material was 5.4 times longer. In the caseof the steel specimens possessing a lesser degree of
cleanliness, the FA-treated material displayed an L10life 3.5 times longer than that of carbonitrided material.
Unfortunately, the testing of the normally quenchedspecimen (with a lesser degree of cleanliness) wassuspended, due to surface peeling developing within
the specimens early on in the testing. It is evident(from the above data) that RCF life can be prolonged
by the refinement of crystalline grains.

Table 3 Metallurgical properties of ball bearing 6206
(0.05mm depth from surface)particle size (μm)

Normally quenched
Carbonitrided
FA treated

Table 2 Test results of point contact fatigue test
Cleanliness

Heat treatment

End face hardness
N
(HRC)
count

Residual

Prior austenite grain(MPa)

Amount ofretained
austenite (%)

+15
-110
-122

7.1
25.5
18.9

HV
stressHeat treatment boundaries Average hardness

L10 life(cycles)

L50 life(cycles)

L10 life ratio

Normally quenched
1462.4
8017 ×104 18648 ×104
1
7
63.0
24656
33974
3.1
Carbonitrided
6
61.6
43244
69031
5.4
FA treated
−1)
−
−
4
62.5
Normally quenched
10
63.6
9018
21653
1.1
Carbonitrided
Lower
10
60.5
30327
55040
3.8
FA treated
1) Life calculation was not possible due to peeling observed on specimens at an early stage of the testing.
Normal

-4-

10.5
9.4
4.4

746
752
733

Improving Rolling Contact Fatigue Life of Bearing Steels Through Grain Refinement

than the carbonitrided-bearing This again shows thatgrain refinement provides a compensatory effect for
the reduction in life that occurs as a result of lower
percentage of retained austenite.

Fig. 3 shows a schematic of the test rig, and Table
4 shows the test conditions.

Loadingcoil spring

4. 2 Life Test on 30206-Type Tapered Bearings
Under Contaminated Lubrication
Conditions

Loading ball bearing 6312
Test bearing 6206

Test bearing
6206

Fig. 3 NTN rolling contact fatigue test rig for ball bearing

The bearings used for the test were normallyquenched, carbonitrided and FA-treated 30206-type
tapered bearings, manufactured from SUJ2 (SAE
52100 EQUIVALENT). Table 6 shows the main
metallurgical constituents from each bearing sample.
Similar to the previous 6206-type ball bearings, theFA-treated bearing sample contains more retained
austenite than the normally quenched bearing but lessthan the carbonitriding.

Table 4 Test condition of 6206 ball bearing undercontaminated lubrication

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

Coupling
Driving pulley

6.86
Load Fr（kN）
Maximum contact stress（GPa）3.2
Rotation speed（min-1）
3000（Inner ring rotation）
Lubricating oil
Turbine oil 56, oil bath, approx. 30ml
Contaminant amount
0.4g/LGas atomized particles :Grain size 100 to 180μm
Contaminant type
Hardness approx. HV800

Heat treatment

Residual Amount ofPrior austenite grain
stressretainedboundaries Average HVhardness(MPa) austenite (%)
particle size (μm)

Normally quenched
Carbonitrided
FA treated

11.8
12.2
5.2

792
763
748

+10
-140
-118

6.0
32.4
23.3

Fig. 4 shows a schematic of the test rig, and Table
7 shows the test conditions.
Table 5 shows the results from the life testing. The
L10 life of the FA-treated bearing was 3.7 times longer
than that of the normally quenched bearing, and 2.1
times longer than that of the carbonitrided sample. It isonce again evident, that RCF life can be prolonged bythis specialized grain refinement method (i.e. FAtreatment). Even under contaminated lubrication
conditions, which can produce indentation-originatedpeeling, it is seen that grain refinement can provide a
positive effect.
Traditionally, it has been known that the greater theamount of retained austenite within steel, the higher
the hardness and the longer the life undercontaminated lubrication 6). However, as shown in
Table 3, the FA-treated bearing possesses a longer
life even though it contains less retained austenite

SpringNormally quenched
Carbonitrided
FA treated

13.1
23.0
48.0

19.4
45.5
87.2

Test bearing

Fig. 4 NTN rolling contact fatigue test rig for tapered
roller bearing
Table 7 Test condition of 30206 tapered roller bearing
under contaminated lubrication
Fr 17.64
Fa 1.5
Maximum contact stress（GPa） 2.5
Rotating speed（min-1） 2000（Inner ring rotation）
Lubricating oil
Turbine oil 56, oil bath, approx. 30ml
Contaminant amount
1.0g/L
Load（kN）

L10 life ratio
N L10 life L50 life
(when the life of normally(h) quenched bearing is 1.0)
count (h)477Support bearing

PulleyTable 5 RCF-life test results of 6206 under contaminatedlubrication
Heat treatment

Testbearing

1.0
1.8
3.7

Contaminant type-5-

Gas atomized particles: 50μm or smaller (80wt%)
100 to 180μm (10wt%) Hardness approx. HV800

NTN TECHNICAL REVIEW No.71（2004）

Table 8 shows the results of the life testing. The L10life of the FA-treated bearing was 4.1 times longer
than that of the normally-quenched bearing, and 2.0
times longer than that of the carbonitrided bearing.

6. Fracture Strength
Bearing ring components (outer diameter:φ60,
inner diameter:φ45, width: 15) were manufactured
from SUJ2 (SAE 52100 EQUIVALENT) and tested toevaluate their static fracture stress.
Investigation on impact strength was also
conducted using JIS3 Charpy impact specimen (Unotched). As shown in Table 10, the carbonitrided
specimen possessed a lower static fracture stress andCharpy impact strength than the normally quenchedspecimen. The strength of the FA-treated specimen
was found to be almost equal to that of the normallyquenched specimen, in spite of having a carbonitrided
layer.

Table 8 The RCF-life test results of 30206 tapered roller
bearing under contaminated lubrication
Heat treatment

N
count

L10 life(h)

L10 life ratio
L50 life(when the life of normally(h)quenched bearing is 1.0)

Normally quenched
Carbonitrided
FA treated

6
6
6

101.2
211.6
415.6

117.3
284.5
464.3

1.0
2.1
4.1

5. Dimensional Change Occurring at
High Temperatures
Table 10 Test results of fracture strength

When bearings operate at high temperatures forextended periods of time, very small dimensionalchanges can occur. One of the main causes for thisdimensional change is an expansion in the material
lattice size due to the transformation of retained
austenite to martensite. Thus, the greater thepercentage of retained austenite within bearing steel,the greater the dimensional change or instability,
making it difficult to use them at higher temperatures.
Since FA-treated bearings possess less retained
austenite than carbonitrided bearings, it is expectedthat they will develop smaller dimensional changes.
To test dimensional stability, the outer rings of6206-type bearings were held at 100˚C and 120˚C
respectively, for 2500 hours, and then measured. The
results are shown in Table 9. It is clear that thedimensional change ratio of the FA-treated specimen
is smaller than that of the carbonitrided specimen.
Specifically, with regards to samples tested at 100˚C,
the dimensional change ratio of the FA-treatedspecimen is approximately 1.3 times less than that of
the normally quenched material. Given the likelihoodthat bearings in the near future will be operating undermore thermally unstable conditions, it will become
extremely important to design units (such as the FAtreated specimen) with high-temperature dimensional
stability.

Heat treatment
Normally quenched
Carbonitrided
FA treated

7. Discussion
As explained in Sections 3 and 4, the FA-treatedmaterial possesses a longer RCF life than theconventional carbonitrided material. This is primarily
due to the delaying of the rolling contact fatigue
mechanism by the grain refinement.
As shown in Fig. 1, the grain size of prior austenite
contained within the FA-treated material is 1/2 times
or smaller than conventional material. However,
because every structure of prior austenite contains
some martensite, it is thought that these structures
have a large influence upon the life and strength of thematerial. So, a FE-SEM/EBSP measuring instrument 7)
was used to measure the distribution of crystal
orientations, to ascertain the difference in martensite
particle size between the carbonitrided and FA-treatedmaterials.
Photo 2 shows an image obtained by FE-SEM/EBSP,
and photo 3 shows the distribution of crystalline
orientation. It is clear (compared to the carbonitrided
material) that the FA-treated material possesses
smaller martensite particles. The areas within 10
degrees of crystal orientation were assumed to be thesame crystal particles, and the diameter of a circle ofthe same area was assumed to be the particle size.
The results reveal the average particle size of the
carbonitrided material as being 0.66 mm, and that ofTable 9 Dimensional change after 2500h-soakHeat treatment

The amount ofDimensional change 1)
retained γ
Temperature: 100˚C Temperature: 120˚C(%)

Normally quenched
7.0
1.0
Carbonitrided
27.7
1.8
FA treated
20.5
1.3
1) Ratio to the normally quenched bearing

Static fracture stress Charpy impact strength(MPa)(J/cm2)
2770
6.70
2330
5.33
2840
6.65

1.0
2.5
1.6

-6-

Improving Rolling Contact Fatigue Life of Bearing Steels Through Grain Refinement

the FA-treated material as being 0.49 mm. Therefore;
it can be considered that extension of RCF life throughgrain refinement is caused by the uniformly reduced
particle size of martensite.

2.40μm

8. Conclusion
By double-quenching through-hardened steel after
carbonitriding, we succeeded in reducing the size ofprior austenite grain size to half that of conventional
steels. Various tests that were conducted demonstrate
the following properties of the grain-refined material.
1） FA-treated material has a RCF life that is
approximately double the RCF life of carbonitrided
material; irrespective of lubrication conditions.
2） FA-treated material developed fewer dimensionalchanges than that of the carbonitrided material(approximately 70%).
3） The static fracture stress and impact strength, thatare reduced as a result of carbonitriding, can berestored.

2.40μmCarbonitrided

FA treated

Photo 2 Image of FE-SEM/EBSP

NTN has begun to utilize this specialized grainrefined material within specific applications. We
believe that as bearings begin to operate underincreasingly hostile environments, the need for thistechnology will expand.

References

Carbonitrided

FA treated

111

111

Miller index
0011）K.Tanaka, TETSU-TO- HAGANE, Vol.79, No.8
(1993), 908(in Japanese)
2）T.Yokota, TETSU-TO- HAGANE, Vol.86, No.7 (2000),
479(in Japanese)3）The Japan Institute of Materials, KINZOKUBINRAN,(2000), 122(in Japanese)
4） H.Murakami, NTN TECHNICAL REVIEW, No.68(2000), 58(in Japanese)5）K.Maeda, NTN TECHNICAL REVIEW, No.65 (1996),
58(in Japanese)
6）N.Tsushima, SAE Technical paper series (1986),
8607257）S.Suzuki, MTERE2, Vol.50, No.7 (2001), 612(inJapanese)

101

Miller index
001

101

Photo 3 Distribution of crystal orientations by
FE-SEM/EBSP

Photos of authors

Chikara OOKI

Kikuo MAEDA

Hirokazu NAKASHIMA

Technical Reserch Dept.
Reserch& Development Center

Technical Reserch Dept.
Reserch& Development Center

Technical Reserch Dept.
Reserch& Development Center-7-

NTN TECHNICAL REVIEW No.71（2004）[ Technical Papers ]Dynamic Analysis of Cage Behavior
in a Cylindrical Roller Bearing

Tomoya SAKAGUCHI*
Kaoru UENO**

It is important to understand time-varying behavior and stress of a cage for an advanced design procedure of rolling
bearings. A dynamic analysis tool is required for that purpose.
Hence, we introduced a commercial software "ADAMS (MSC. Software)" for virtual prototyping of a mechanicalsystem. The environment provides a set of high-performance numerical integration solvers to resolve equations of
motion as well as high-level visualization functions of the numerical results.
In this report, as a milestone for the development of dynamic analysis tools of various types of rolling bearings, wedeveloped a code on ADAMS that can simulate real-time behavior of a cylindrical roller bearing assuming the motion of
all the bearing elements is planar. Also, with regard to the cage behavior, we made a comparison between numericaland experimental results. The results show that interaction forces between the cage and the rollers traveling aroundthe exit of the load zone have a major influence on the cage motion.

and experimental results.
The results show that interaction forces between thecage and the rollers traveling around the exit of theload zone have a major influence on the cage motion.

1. Introduction
It is important to understand time-varying behavior
and stress of a cage for an advanced design
procedure of rolling bearings. A dynamic analysis tool
is required for that purpose.
Hence, we introduced commercial software"ADAMS 1)" for virtual prototyping of a mechanicalsystem. The environment provides a set of highperformance numerical integration solvers to resolve
equations of motion as well as high-level visualization
functions of the numerical results.
In this report, as a milestone for the development ofdynamic analysis tools of various types of rolling
bearings, we developed a code on ADAMS that cansimulate real-time behavior of a cylindrical roller
bearing assuming the motion of all the bearing
elements is planar. Also, with regard to the cage
behavior, we made a comparison between numerical

2. Symbols
an：Amplitude of n -th order waviness of cageroundness, mb：Hertz contact half-width, mbn：Phase of n-th order waviness of cageroundness, radCr：Basic dynamic radial load rating, N

D：Non-dimensional
Deborah number＿［＝η0 eαP u/Gb］￣

E：Young's modulus, Pa

E'：Equivalent Young's modulus, Pa

FEHLr ：Shearing force by viscosity of EHL film, N**Technical Research Dept., Research & Development Center**Industrial Engineering Department, Industrial Sales Headquarters-8-

Dynamic Analysis of Cage Behavior in a Cylindrical Roller Bearing

Fhydro ：Viscous shearing force of oil film at cage guide, Nhydrodynamic lubrication

Fpx ：Rolling directional component of EHL film
pressure, N
FT

μbd ：Traction coefficient under boundary lubrication
μhd ：Traction coefficient under hydrodynamic lubrication
μr ：Traction coefficient at contact area

：Tangential force at contact area, N

Fy,Fz ：Force acting on cage at cage guide, Nν：Poisson's ratioΣ：Dimensionless shearing velocity of lubricant［＝η0 e αP s/（τ0 hc）］

f

：Frequency, Hz

G：Dimensionless material parameter［＝α0 E'］h：Oil film thickness, m

τ0：Characteristic stress of lubricant, Pa

hc：Central oil film thickness, mτr：Shearing stress by viscosity of EHL film
during pure rolling, Pa―

hc, iso ：Isothermal central film thickness, mk'：Thermal conductivity of lubricant, W/mKτs

：Shearing stress by slip of EHL film, Pa

lbe：Roller effective width, m

φ：Angle, radlco：Effective width (one side) of cage guide face, m

φc：Angle at cage center, radφr：Cage revolution angle, radLT＿：Thermal load factor［＝η0βu2/k'］

φTH ：Oil film thickness correction factor by
shearing heat of EHL oil film＿

：Mean pressure of Hertzian contact, Pa
P
PHZ ：Maximum pressure of Hertzian contact, Pa
Q：Normal force at contact area, N

φTR ：Correction factor by shearing heat of EHL
rolling viscous resistanceR：Radius of curvature, mω

Rco ：Radius of cage outer diameter, m(Average value if an upper bar is present)
Re：Equivalent radius, mRg：Radius of outer ring inner diameter, m＿S：Mean dimensionless shearing stress ofcontact areas［＝τs/τ0］s：Slip ratio［＝｜ub−ur｜/u］

t：Time, sU/ E'R）］：Dimensionless velocity［＝μ0 u（u：Surface velocity, m/s＿：Revolution angle of each element, rad/s

Subscript
b
c
i
o
r

＿

：Roller：Cage：Inner ring
：Outer ring
：Raceway＿u：Mean velocity, m/s［＝0.5×｜ub＋ur｜］

W：Dimensionless load parameter［＝Q/（E'Relbe）］

Xc：Dimensionless length of EHL contact area［＝（D/Σ）
sin h-1Σ］

3. Analysis Method for Cage Behavior
Fig. 1 shows a schematic of the test bearing and its
coordinate system. All the coordinates are expressedin the right-handed Cartesian coordinate system.
While the outer ring is fixed in space, the inner ringrotates clockwise about the -X axis and the radial load
is in the +Z direction. Gravity is in the -Z direction.
Principal assumptions are given below.

yc ,zc ：Geometric center position of cage, m
y0 ,z0 ：Geometric center position of outer ring, mNz：Number of rolling elementsα：Viscosity pressure coefficient, 1/Pa

α0：Pressure viscosity index of lubricating oil
under normal pressure, 1/Pa

β

：Viscosity temperature-rise coefficient, 1/K

δ：Geometric interaction amount, mε：Eccentricity ratio for journal bearing modelη0：Viscosity under normal temperature andpressure, Pa・sθ

：Angle, radΛ：Film parameter (Ratio to oil film parameter,
combined roughness)

1. The inner ring and cage can move on the radialplane (Y-Z plane in Fig. 1) only.
2. Contact force in geometric interaction areas
conforms to the elastic contact theory.
3. Traction characteristic conforms to Formula 2) set by
Mr. Muraki et.al..
4. Rolling viscous resistance 3) by EHL film exists inthe contact area between the rollers and raceway.
5. The interaction model at the cage guide isexpressed using the theory of short-width journalbearing with squeeze effect and cage's outer
diameter roundness taken into account, andGümbel's boundary condition is adopted.
6. Gravity is taken into account.

Λbd ：Maximum film parameter under boundary
lubrication
Λhd ：Minimum film parameter under-9-

NTN TECHNICAL REVIEW No.71（2004）

then calculate the film thickness using the PanHamrock's formula for isothermal central film
thickness of EHL in line contact 4) and GhoshPandey's thermal correction factor equation 5).

Displacement
sensor

Z

Z

X

0.9
Q 0.9
8
δi,o ＝ 0.39 ―― ――――………………………（1）E'
lbe 0.8

（ ）

Y

O

O

Direction ofrotation-0.166
0.692
0.47
U
G Re （2）hc ＝φTH hc,iso ＝φT・2.922 W

where,｛-1

Re ＝

Radial load

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

Fig.1 Schematic of test bearing and its coordinate system［

E' ＝ 2

1 −νr2
1 −νb2＋
Eb
Er………（3）-1］………………（4）-1

φTH ＝［1 ＋ 1.6 W 0.152 LT 0.379（1 ＋ 3.96s 0.96）］（5）As the initial condition for numerical integration
solvers of equations of motion, it is assumed that therollers and cage rotate on their own axis or rotate
around the other at mechanics theoretical speed. It isalso assumed that the cage is centered at the outer
raceway and the rollers are located on the pitch circle
line as well as the center of the cage pocket.
Rotational displacement was forced to the inner ring at
a constant angular velocity.
Physical observation time was set to 0.3 seconds
and data recording was performed at 0.5-ms intervals.
Calculation model for each interaction area isshown below.

4）From this calculated film thickness, obtain traction
coefficient (μhd) for hydrodynamic lubrication
using Muraki-Kimura's simplified equations2)
shown below.
Xc ＞ 2：S ＝ D/Σ ……………………………（6）
Xc ＜ 2：S ＝ sinh -1 Σ
｛1 −（D/4Σ）sinh -1Σ｝ …（7）μhd ＝τ0 S/P …………………………………（8）Where, Xc ＝（D/Σ）sinh-1 Σ indicates thedimensionless length of EHL contact area.

3.1 Interaction Force between Roller Rolling
Contact Surface and Raceway
The model of interaction force between roller rolling
contact surface and raceway is based on the following
assumptions.
1）Because of the relative positions of the rollers andraceway given by time t, normal force by contact
and tangential force and rolling viscous resistance 3)
by EHL film will occur between the two objects
only when geometrical interaction occurs. For
analysis of dynamics, squeeze effect (dependantof speed) by EHL film must be taken into account.
However, this is not considered since doing so will
require accurate EHL calculation that necessitates
tremendous cost.
2）If the EHL film is so thin that it will tear, friction bysolid contact must be taken into account for thetangential force.
3）Consider that line contact of the roller's effective
length (lbe) occurs if geometrical interaction
amount (δ) is given, and calculate normal force Q
using Palmgren's simplified equation. Determine
non-dimensional load variable W based on Q, andFor equation (7), it is necessary to consider
temperature rise caused by self heating of the film,
therefore, convergence calculation is required.
From this, the traction coefficient of the EHL film
can be determined.
As shown in assumption 2), the traction
coefficient (μr) for the entire set of all lubrication
regimes can be approximated by equation (9), with
influences of boundary lubrication and mixed
lubrication taken into account.
μbd
μbd −μhd6
μr ＝（Λ−Λhd）＋μhd（Λbd −Λhd）
μhd

｛

｛

if Λ＜Λbd｝

if Λbd ≦Λ＜Λhd

if Λhd ≦Λ……………………………（9）The tangential force (FT) with μbd = 0.1，Λbd = 0.06，
Λhd = 3.0 can be calculated by equation (10).
｜FT｜＝μr Q …………………………………（10）-10-Dynamic Analysis of Cage Behavior in a Cylindrical Roller Bearing

in the same way as 3.1, "Interaction Forcebetween Roller Rolling Contact Surface andRaceway".
4）From this calculated film thickness, obtain traction
coefficient using Muraki-Kimura's simplified
equations (6) to (8) 2).
5）If the center of the rollers comes out of the obliqueline area (see Fig. 2) enclosed by the pocket
center and pocket ends, only the interaction force
caused by solid contact between the roller andpocket edge must be taken into account.

5）For the EHL rolling viscous resistance, use ZhouHoeprich's equation (11) 3). The direction of FEHLr
is the same for both objects, and is opposite to thecomposite velocity vector on the contact surface ofthe two objects. For the rotating two objects, theforce (Fpx ) caused by the pressure component in
the EHL film rolling direction must be taken into
account 3).
FEHLr ＝φTR

0.648
0.246
29.2Re lbe（GU） W
α0Provided that, φTR ＝

Fpxb,r ＝−……（11）

3.3 Interaction Force between Cage and Outer
Ring (Guiding surface)1 − 13.2（PHZ − / E'）LT 0.42
1 ＋ 0.213（1 ＋ 2.23s

0.83

）LT

0.64

Since the outer ring guide type cage is used, only
the interaction force between the cage and outer ring's
rib inner diameter can be taken into account. The
assumptions are given below.
1）Since the guide's width is smaller than the
diameter, a short-width journal bearing theory is
used.
2）The bearing is placed under a hydrodynamic
lubrication for an isoviscous-rigid regime and alaminar flow, and the squeeze effect and cage's
outer diameter roundness must be taken into
account. The cage's outer diameter roundness isexpressed by equation (13).

2Re
FEHLr …………………………（12）Rb,r3.2 Interaction Force between Rollers and Cage
Pocket
The interaction force model is almost the same as
the one between roller rolling contact surface andraceway. However, since the traction force isdominant due to large slip ratio, the EHL rollingviscous resistancet 3) can be ignored. The model forinteraction force between rollers and cage pocket isgiven below.

m

Rco（θ）＝ Rco ＋Σ an cos（nθ＋ bn） ………（13）1）Because of the relative positions of the rollers andcage pocket given by time t , normal force bycontact and tangential force by EHL film will occur
between the two objects only when geometricalinteraction occurs. Squeeze effect for theseobjects is ignored.
2）If the EHL film is so thin that it will tear, friction bysolid contact must be taken into account for thetangential force.
3）Consider that line contact of the roller's effective
length (lbe) occurs if geometrical interaction
amount δ is given, and calculate normal force Q

n=13）Only positive pressures among those obtained by
Reynolds equation are taken into account(Gümbel's boundary condition).
4）If the clearance is infinitely small in the case ofjournal bearing, an infinitely large interaction force
will occur theoretically. Thus, the interaction force
caused by direct contact between the cage andouter ring can be ignored.

Pocket center
Pocket pillar

Fig. 2 Cross section of cage pocket-11-

NTN TECHNICAL REVIEW No.71（2004）

Equation (16) indicates that negative pressures are
ignored as cavitation condition.
The viscous frictional force (Fhydro) acting along the
cage's outer diameter surface is expressed byequation (20) as a sum of two equivalent journalbearings. "Rco×Fhydro" acts as a moment to preventthe cage from rotating on its axis.

Equations for the short-width journal bearing basedon the above assumptions are given below. The
coordinate system for this model is shown in Fig. 3.
The force given on the cage by the film is expressedby a sum of two equivalent journal bearings, and is
obtained by equation (14) by y-z plane.

zh｜Fhydro｜＝η0ωc Rco2 lco

4π

Rg − Rco

1 −ε2Outside of cageφ

θwhere, ε indicates the eccentricity ratio as given inequation (21).

Inside of outer ring rib
c

φr

φ

y

Oc
O

…………（20）

ε＝
Rc

o

2
2
（yc − yo）＋（zc − zo）

Rg − Rco…………………（21）Rg

4. Cage Behavior Measuring Method
Attach the test bearing to the horizontal type spindlethat rotates clockwise about the -X axis, and exert +Z
radial load on the outer ring. Gravity is in -Z direction.
Table 1 shows the specifications of the test bearing
and running conditions. The cage is made of brass,and is guided by the inner diameter surface of theouter ring's rib as shown in Fig. 1. The cage is short in
the arc length, but has a cylindrical pocket surface(see Fig. 2).
Measurement of cage behavior is carried out using
two eddy-current displacement gauges for both Y andZ directions. The reason for using two gauges for the
X direction is to check the cage for absence of conicaloscillation.

Fig. 3 Journal bearing coordinate system for outer ringland riding cage

cosφ
Fy
2π
3
＝η0 Rco lco ∫0 （g φ）sinφ d φ ……（14）
Fz

［ ］

［ ］

1
h3

（
g φ）＝

｛

（
g φ）＝

［

ωc

∂h
∂h
＋2
∂φ
∂t

］

（
g φ） if （g φ）＜ 00…………（15）…………（16）otherwise

Table 1 Test bearing and operating conditions

h ＝（Rg −Rco（φ−φr））−（yc − yo）cosφ−（zc − zo）sinφ

Bearing type
（Bore × O. D. × Width, mm）

NU2310G1（φ 50 ×φ 110 × 40）…………（17）

Number of rollers
Basic static load rating C0r, N
Cage typeRadial internal clearance, μ m
Cage guide clearance, mm

12
131 000
Machined, Outer ring land riding
40, 5
0.445

LubricantNo-additive turbine oil VG56,
Air-oil lubrication

Rotational speed, min-1
Radial load Fr, N
Temperature of outer ring at O.D. ˚C1000, 3000, 5000
980, 4900
35 ± 3m∂h＝Σ nan sin n（φ−φr）＋ bn ＋（yc − yo）sinφ∂φ n=2［｛｝］−（zc − zo）cosφ ………………………（18）∂h-nωc an sin n（φ−ωct）＋ bn
＝
∂t Σ
n=2
・ ・
・ ・
−（yc − yo）cosφ−（zc − zo）sinφ ………（19）
m

［

｛

｝］

-12-

Dynamic Analysis of Cage Behavior in a Cylindrical Roller Bearing

5. Behavior of Cage Center

1
Dimensionless cage center position Z

5.1 Analysis Results
Fig. 4 shows an example of oscillation occurring at
the cage center during analysis. The displacement
amount at the cage center shown in Fig. 4 was madedimensionless using the radius clearance of the cage
guide. The initial cage center position was the origin ofFig. 4. The cage center moves considerably
immediately after the start of calculation, but
eventually reaches steady state. The steady state ofthis cage accompanies minute vibration as shown in
Figs. 4 and 5.
Fig. 6 shows the result of frequency analysis
conducted on the waveform shown in Fig. 5, startingfrom the 0.08-second point following the start ofcalculation, where the behavior of the cage center has
reached the steady state. The amplitude of the motion
is large at the cage's self-rotation frequency fc, its fourfold and Nz -fold frequencies. Since the amplitude atNz fc is especially large compared to those at 8 fc and
16 fc, it can be said that the vibration frequency of thecage center is largely influenced by the roller's rotating
frequency around the outer ring.

5000min-1
0.5
3000min-1

0-0.5

5000min-1
3000min-1
1000min-1

1000min-1

-1
-1

-0.5

0

0.5

Dimensionless cage center position Y

Fig. 4 Numerical results of cage center behavior(Fr : 4900N, Radial clearance: 40 μm)

Cage center position Z, mm

0.15
0.1
0.05
0
-0.05
-0.1
0

0.05

0.1

0.15

0.2

0.25

0.3

Time, sAmplitude of cage center motion, mm

Fig. 5 Numerical result of cage center position along Z axis (Fr : 4900N, radial clearance: 40μm, 3000min-1)
fc

4 fc

Nz fc

0.01
0.001
0.00010.00001
0.000001
0

100

200

300

400

500

600

700

800

900

1000

Frequency, Hz

Fig. 6 Frequency analysis of cage vertical steady motion(Fr : 4900 N, radial clearance: 40μm, 3000min-1, 841 points, hanning window, fc : 20Hz, Nz fc : 240Hz)

-13-

1

NTN TECHNICAL REVIEW No.71（2004）

5.2 Measurement Results
Fig. 7 shows examples of waveforms obtained by
the four displacement gauges. Steady motion withminute vibration was observed at the cage center. The
signals from the displacement gauge pair for X
direction show that there is no phase shift, and no
conical oscillation of the cage was found.
Fig. 8 shows the frequency analysis results for theSensorpositioncage center behavior. The bearing running conditions
are the same as those for Fig. 6. Like Fig. 6, theamplitude of motion is large at the cage's self-rotation
frequency fc and Nz fc (roller's orbital rotating
frequency around the outer ring), and the frequency ofminute vibration at the steady state is well resembled
between analysis and measurement data.

0.04 mm/div

Y direction
at -X side

Z direction
at -X side

Y direction
at +X side

Z direction
at +X side
0.1 s/div
One revolution cage

Fig. 7 Experimental results of cage behavior (Fr : 4900 N, radial clearance: 40μm, 5000min-1)

fc 2 fc 4 fc

Nz fc

2Nz fc-20.0
PWR SP
MAG-120.00.0000［Hz］

han

1.0000K

Fig. 8 Frequency analysis of cage vertical motion measured by the top sensor(Fr : 4900N, radial clearance: 40μm, 3000min-1, 2048 words, hanning window, fc : 20Hz, Nz fc : 240Hz)-14-

Dynamic Analysis of Cage Behavior in a Cylindrical Roller Bearing

5.3 Comparison of Cage Steady Position

6. History of Interaction Force between
Roller and Cage

Fig. 9 shows a summary of the analysis andexperimental results regarding the center position ofminute vibration at the cage center during steadystate. In Fig. 9 a), the angle from the origin to thecage center is slightly different between analysis andexperiment results, but the same qualitative trend ofrotating speed increase is shown in both results. In
Fig. 9 b), no experimental results are given in thecase of 1000min-1, since the cage did not stabilize. In
the case of 4900N load, variation direction of the cage
center position for rotational speed rise does notmatch that of the analysis result, but the center
position itself is in close agreement. On the whole,validity of analysis can be confirmed for both trend
and position of cage center variation.

Fig. 10 shows an example of variation in roller's
self-rotating slip ratio, roller's circumference position
inside the pocket, roller's circumferential contact force
to the pocket and roller's contact force to the innerring. Signs (+, -) given to the roller's circumferential
contact force to the pocket mean as follows: "+"indicates the direction in which the cage isaccelerated, and "-" indicates the direction in which
the cage is decelerated. From the variation of thecontact force to the inner ring, the time for the roller topass the load zone can be found. The roller generates
a maximum of 22% slippage in the non-load zone andthe roller circumferential position is located toward theback of the pocket orbital motion. The roller begins toaccelerate rapidly at the entrance of load zone andcontinues until it reaches the theoretical speed. In thelatter half of the load zone, the roller comes into
contact with the pocket surface as shown in Fig. 11.
The reason why the slip ratio is large in the non-loadzone is that EHL rolling viscous resistance is also
taken into account like it is in the load zone, despite
the fact that only centrifugal force, which is relatively
small, is exerted at the area contacting with the outerring raceway. Thus, it appears that the rolling viscous
resistance is taken into account more than it should be.
As a result, the force that causes the roller toaccelerate the cage in the second half of the loadzone is added in the direction shown in Fig. 11. Since
that force is supported at the cage guide, the cage
center position stabilizes at the position shown in Fig. 9.
Fig. 9 shows that the motion variation in theanalysis result is small, but the cage tends to move in
Z direction due to reduction of the radial clearance or
increase of load. This is due to the change in roller
contact force direction in the second half of load zone
caused by extension of the load zone.
Furthermore, the cage moves in Z direction when
the rotating speed increases. The reason is given
below. First, resisting moment on the cage guidesurface increases due to increased rotating speed.
The roller contact force to the pocket in the second
half of load zone then increases in order to cancel thisresisting moment. Since this contact force acts as atranslation force for the cage, the force to the upper
left corner in Fig. 9 increases, causing the cage tomove in Z direction.

Dimensionless cage center position Z

1Analysis (980N)
Experiment (980N)
Analysis (4900N)
Experiment (4900N)

0.50Speed up (1000⇒3000⇒5000min-1)
-0.5

-1
-1

-0.5

0

0.5

1

Dimensionless cage center position Ya) Radial clearance : 40μm

Dimensionless cage center position Z

1
Speed up (3000⇒5000min-1)
Analysis (980N)
Experiment (980N)
Analysis (4900N)
Experiment (4900N)

0.50Speed up (1000⇒3000⇒5000min-1)
-0.5

-1
-1

-0.5

0

0.5

1

Dimensionless cage center position Yb) Radial clearance : 5μm

Fig. 9 Comparison of numerical and experimental results
for cage center steady position-15-

NTN TECHNICAL REVIEW No.71（2004）

Roller position in pocket
Roller force to inner / 100

0.3

60

0.2

50

0.1

40

0

30

-0.1

20

-0.2

10

-0.3

0
-10

-0.4
Load zone-0.5
0.2

Contact force, N

Slip ratio and roller position, mm

Roller slip ratio
Roller force to pocket

0.22

Load zone

0.24

0.26

0.28

-20
0.3

Time, sFig. 10 Numerical results of roller behavior and roller/cage interaction (Fr : 4900N, radial clearance: 5μm, 3000min-1)

Blue arrow:
Outer force to roller

Roller force to pocket

Inspected roller
located among load zone

Light red arrow:
Inner force to roller

Fig. 11 Graphic example of interaction forces (Fr : 4900N, radial clearance: 5μm, 3000min-1）-16-

Dynamic Analysis of Cage Behavior in a Cylindrical Roller Bearing

7. Conclusion
Analysis of real-time behavior of a cylindrical roller
bearing on a radial plane was carried out using
commercial software designed for virtual prototypingof a mechanical system. As a result of verification of
the experimental results of the behavior of the cage
center, validity was confirmed. It was found that roller
contact force in the second half of load zone in the
direction in which the cage's self-rotation isaccelerated was large and the cage center position isdetermined by this force.
This report limits to real-time analysis under staticload. However, analysis for desired running conditions
such as fluctuating load is also possible. With the
purpose of drastic improvement in bearing design
technology, we plan to develop a code on thissoftware that can simulate the real-time threedimensional behavior of a cylindrical roller bearing.

References
1) MSC. Software, HP Address：
http://www.adams.co.jp/（2003.5.7）
2) M. Muraki, Y. Kimura: J.JSME, 28,10(1983)753760.(in Japanese)
3) R. S. Zhou, M. R. Hoeprich: Trans. ASME, J. Trib,
113, 7(1991) 590.
4) P. Pan, B.J. Hamrock: Simple Formulae for
Performance Parameters Used in
Elastohydrodynamically Line Contacts, Trans.
ASME, J. Trib., 111, 2(1989) 246-251.
5) M.K. Ghosh, R.K. Pandey: Thermal
Elastohydrodynamic Lubrication of Heavily Loaded
Line Contacts-An Efficient Inlet Zone Analysis, Trans.
ASME, J. Trib., 120, (1998) 119-125.

Photos of authors

Tomoya SAKAGUCHI

Kaoru UENO

Technical Reserch Dept.
Reserch& Development CenterIndustrial Engineering Department,
Industrial Sales Headquarters-17-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Precision Bearings "ULTAGE" Series for Machine Tools

Futoshi KOSUGI*

Recently, machine tool manufacturers have placed increasingly demanding performance specifications on the
bearings used in their products. Machine Tool Bearings are required for high speed, high efficiency and highprecision. In recent years, dry machining technology, which does not use oil in the machining process, is also
becoming an important factor in the improvement of working environment.
NTN has been pursuing high dmn values and high precision of bearings for the main spindles of machine tools andconcentrating on technology for improvement of working environment. This paper introduces ULTAGE Series formachine tools that have been recently developed not only for high speed and high precision but for harmony with theenvironment.

1. Introduction

2. ULTAGE Series

Machine tools, including machining centers, arelooking for higher speed, higher efficiency and higher
precision. In addition, as the development of dry
machining technology that does not use oil during themachining process suggests, more attention is being
paid to the working environment.
With regard to bearings designed for the main
spindles of machine tools, NTN has been pursuing
higher dmn values, higher precision and lower NRRO
(Non-Repeatable RunOut). We are also paying
attention to eco-consciousness.
This paper introduces the "ULTAGE Series",
developed as ECO series for machine tools, with theconcept of "harmony with the environment" in addition
to high speed and high precision features. "ULTAGE"
is a combination of the words "ULTIMATE" and"STAGE", to express the NTN's concept of pursuing
the ultimate precision rolling bearings for machine
tools.

The ULTAGE Series offers a total of 10 types ofbearings, including angular contact ball bearings andcylindrical roller bearings. (See Table 1)
For angular contact ball bearings, a total of eighttypes are available. They include "70U/79U" typedeveloped from the standard angular contact ballbearings; "HSE" type with improved abrasion
resistance and seizure resistance in addition to highspeed capability; "HSF" type which achieves higher
speed and limited temperature rise; two sealed types(BNS, 70/79AD, CD) designed for grease lubricatedapplications; the eco-conscious "HSL , HSFL" types
designed for air-oil lubricated applications, with
reduced noise level and low air/oil consumption; theball screw support type "2A-BST", and the light
contact sealed type "2A-BST LXL".
For cylindrical roller bearings, two types areavailable; "N10HSR type" featuring high-speed
operation under both air-oil and grease lubrication,and the eco-conscious "N10HSL type" designed for*Industrial Sales Headquarters Industrial Engineering Department-18-

Precision Bearings "ULTAGE" Series for Machine Tools

air-oil lubricated applications only, featuring reduced
noise level and low air/oil consumption.

In this paper, five types (sealed and eco-conscious
types) in the ULTAGE Series are introduced.

For spindles
For spindles

For ball screw support

Angular contact ball bearing
Cylindrical roller bearing

ULTAGE series

Table 1 ULTAGE series[Standard]
70U type , 79U type(Contact angle 15˚, 25˚, 30˚)

Optimized interior structure and resin cage help positively
inhibit temperature rise.

[High-speed]
5S-2LA-HSE type(Contact angle 15˚, 20˚, 25˚)

Adoption of special material and surface treatment, withdrastically improved abrasion resistance and seizure
resistance. Features improved rigidity and reliability in addition
to the high speed of the conventional "HSB0CAEX1" ultra
high-speed angular contact ball bearing.

[Super high-speed]
5S-2LA-HSF type(Contact angle 25˚)

Maintaining the advantages of HSE type, this type has small
diameter ceramic balls to achieve higher speed and limited
temperature rise.

[Eco-conscious]
5S-2LA-HSL type(Contact angle 15˚, 20˚, 25˚)
5S-2LA-HSFL type(Contact angle 25˚)

Adoption of special material and surface modification, with drastically
improved abrasion resistance and seizure resistance.
Designed only for air-oil lubrication, with circumferential grooves on
the inner ring of the HSE type and eco-conscious nozzle. Features
reduced noise and air/oil consumption in addition to the high speedof the HSE type. It is applicable also to HSF type.

[High-speed]
5S-2LA-BNS type(Contact angle 15˚, 20˚, 25˚)

Adoption of special material and surface heat treatment , withdrastically improved abrasion resistance and seizure
resistance. Designed only for grease lubrication, with
optimized internal design, grease reservoirs, special grease
and non-contact seals.

[Standard]
70CD type, 79CD type(Contact angle 15˚)
70AD type, 79AD type(Contact angle 25˚)Angular contact ball bearing with optimized internal structure,
non-contact rubber seals on both sides, and long-life special
grease. Bearing with ceramic balls is also available.

[Open]
2A-BST type(Contact angle 60˚)Open type bearing with longer rolling contact fatigue life byadditional surface heat treated rings.

[Light contact sealed]
2A-BST LXL type(Contact angle 60˚)

Features longer rolling contact fatigue life by additionalsurface heat treated rings and drastically reduced fretting
wear by employing special grease.
Provides improved dust resistance and grease-retaining
capability by low-torque and light contact seals.

N10HSR typeFeatures optimized internal structure for high speed capability
with low temperature rise. Provides higher limiting speed thanthe conventional "N10HS" high-speed cylindrical roller bearing.

N10HSL type

Designed only for air-oil lubricated applications, withcircumferential grooves on the inner ring of the N10HSR typeand an eco-conscious nozzle. Features reduced noise andlow air/oil consumption in addition to the high speed capabilityof the N10HSR type.

-19-

NTN TECHNICAL REVIEW No.71（2004）

2. 1 Bearings for Main Spindles
2. 1. 1 Grease-lubricated Sealed Angular
Contact Ball Bearing (BNS Type)

Orange seal

From the viewpoint of environmental issues, greaseis the most suitable lubrication method for bearings
used for main spindles. When using bearings undergrease lubrication, no external lubrication unit isrequired, therefore assembly and maintenance areeasy. Additionally, because the bearings are prelubricated with a suitable amount of grease, theamount of oil mist is very low.
However, when using high-speed angular contact
ball bearings, the lubrication life will be a major factor.
NTN evaluated the long life grease from various
viewpoints, and developed "Grease-lubricated Sealed
Angular Contact Ball Bearing (BNS Type)" 1) (Fig. 1)
to be used for high-speed spindles of φ50 or larger
diameter. Then, NTN introduced at JIMTOF2000(Japan International Machine Tool Fair).
For the BNS type, special material and surface heat
treatment are adopted for rings to extend the rolling
contact fatigue life. Grease is retained and supplied to
the raceway surface by the seals provided on bothsides of the bearing. Grease reservoirs set directly
adjacent to the outer ring raceway, in order to extend

Black seal

Fig. 1
Grease-lubricated Sealed Angular Contact Ball Bearing

the grease life during high-speed operation.
Fig. 2 illustrates the effect of grease reservoirs forbearing life, and Table 2 shows the results ofendurance test
From the test results shown in Fig. 2, it was foundthat the life of the grease-lubricated sealed angularcontact ball bearing is 7.4 times longer than theconventional type.
The endurance test was conducted on the BNS type(5S-2LA-BNS020: φ100×φ150×24) with seals on
both sides, and HSE type (5S-2LA-HSE020C without
seals: φ100×φ150×24), with "0N" fixed position[Test Conditions]
Grease reservoir

7.4 timesWith
greasereservoirs
Without
greasereservoirs

260h[With grease reservoirs]
5S-2LA-BNS020(φ100×φ150×24)(Without seal)[Without grease reservoirs]
5S-2LA-HSE020(φ100×φ150×24)(Without seal)

35h

Grease: NBU15
Rotating speed : 13,000 min-1
Jacket cooling : Provided

Fig. 2 Effect of grease reservoirs on bearing life
Table 2 Endurance test results[Test Conditions]
Bearing

Spindledirection

HSE type
Horizontal(Without seal)

Duration , h
0

5000

10000

15000

20000

[HSE type]
5S-2LA-HSE020C (φ100×φ150×24)[BNS type]
5S-2LA-BNS020LLB(φ100×φ150×24)

650h

BNS type

Horizontal

20000hBNS type

Vertical

20000h-20-

Grease: MP-1
Rotating speed : 11,000 min-1
Preload amount : 0N after mounted(Fixed position preloading)Jacket cooling : Provided

Precision Bearings "ULTAGE" Series for Machine Tools

preload (after mounted) and at a rotating speed of
11,000 min-1 (dmn value: 1,400,000). The HSE typeshowed a life of 1,000 hours or shorter, while the BNS
type showed a life of 20,000 hours (the test wassuspended after 20,000 hours).
This result suggests that air-oil lubricated spindles
can be replaced by grease-lubricated spindles in thehigh-speed range up to 1,400,000 (dmn value).

This bearing has a load capacity equal to thestandard angular contact ball bearing, and offers two
types: CD type (contact angle 15˚) and AD type(contact angle 25˚). It is developed based on thedesign concept of the BNS type, which features noncontact seals on both sides, long-life special grease
and optimized internal structure for reduction of heatgeneration. Bearing with ceramic balls is alsoavailable.
Fig 4 shows the results of high-speed operation
test. The test was conducted on the 70CD (7006CD:
φ30×φ55×13) and 70AD type (7006AD: φ30×φ
55×13), with mounted preload of 180N and 250N,
respectively. The bearings were mounted with DB
arrangement . Both 70CD and 70AD types ran stableat rotating speeds of up to 25,000 min-1 (dmn value:
1,100,000).
With 70/79 AD and CD types, black and orange
seals were provided at the front and rear sides
respectively, like the BNS type, to facilitate
confirmation of bearing mounting orientation duringassembly work.

In addition, since the BNS type is a pre-lubricatedsealed type, there is no need to fill grease and toclean the bearings during spindle assembly. This may
not only simplify handling procedure and reduce
assembly labor, but also improve the working
environment (e.g. elimination of cleaning oil disposal).
Furthermore, an improvement was made on theseal after JIMTOF2000. Seal on the back side was
replaced by the orange seals (previously, black seals
were used on both sides). This facilitates confirmationof bearing direction during assembly work: for DB
mounting, the orange seals face each other, and for
DF mounting, the black seals face each other.

2. 1. 2 Grease-lubricated Sealed Angular
Contact Ball Bearing (70/79 AD, CD Type)

Orange seal

Black seal

For spindles of φ50 or less inner diameter, such as
rotating tool spindles and small-size spindles, which
are used in low to medium speed ranges, wedeveloped the "Grease-lubricated Sealed Angular
Contact Ball Bearing (70/79 AD, CD Type)" (Fig. 3) 2)
and introduced at JIMTOF2002.
Fig. 3
Grease-lubricated Sealed Angular Contact Ball Bearing

Table 3 Bearing arrangement aided by seal color

Orange seal＋

Outer ring temperature rise ˚C

DB set [Rear-side]
Orange seal

40

70CD type70AD type

35
30
25
20
15
10
5
0

DF set [Front-side]

0

5000

10000

15000

20000

Rotating speed min-1
Black seal＋

Black seal

[Test Conditions][70CD type]
7006CDLLB (φ30×φ55×13)
Preload amount :180N after mounted
Jacket cooling:Not provided[70AD type]
7006ADLLB (φ30×φ55×13)
Preload amount :250N after mounted
Jacket cooling:Not provided

Fig. 4 High-speed test results

-21-

25000

NTN TECHNICAL REVIEW No.71（2004）

circumferential groove at the exit of the nozzle, andwhen collected oil enters the inner section of the
bearing at once, it causes the bearing temperature tofluctuate. This issue was solved by removing thecircumferential groove from the nozzle and instead
providing circumferential groove on the outer face ofthe bearing inner ring. 4) 5)
Fig. 5 shows the bearing design of HSE(HSF) type ,SF type ,and HSL(HSFL) type.
Fig. 6 shows the correlation between air supply rate
and outer ring temperature measured for the HSF
type, SF type and HSFL type (5S-2LA-HSFL020: φ
100×φ150×24), with a constant spring preload(2.5kN). The test was carried out under fixed
conditions: rotating speed of 21,000 min-1 (dmn value:2,650,000) and oil supply rate of 0.03mL/5min.
With the HSF and SF types, the outer ringtemperature rose suddenly at an air supply rate of
22.5NL/min and 15NL/min respectively. However, with
the HSFL type, operation was still possible at much
lower air supply rate of 10NL/min.
Fig. 7 shows change of outer ring temperature at a
fixed rotating speed (21,000 min-1) and at different oil
supply intervals (oil supply rate: 0.03mL/shot). With
the HSFL type, stable operation was possible in the oil
supply interval range of 2 to 21 min. With the HSF

2. 1. 3 Eco- conscious Air-Oil LubricatedAngular Contact Ball Bearing (HSL ,HSFL Type)
Advantages of grease lubrication with regard toenvironmental issues were discussed previously.
However, sufficient supply of lubricating oil to the
raceway is difficult in high-speed operation, thereby,
the life of lubricating oil becomes an issue. Because ofthis, air and oil lubrication is normally used in highspeed applications. However, reduction of air/oil
consumption and noise level are constantrequirements.
Based on the "Low-Noise Angular Contact Ball
Bearing (SF Type)" 1), which was introduced at
JIMTOF2000, we continued to make further
improvements and have eventually developed the"Eco-conscious Air-Oil Lubricated Angular Contact
Ball Bearing (HSL , HSFL Type)" 3), featuring not only
high-speed capability and low noise, but also reduced
air and oil consumption, which was then introduced at
JIMTOF2002.
Even with the SF type (low-noise angular contact
ball bearing), oil was supplied efficiently to the insideof the bearing and reduction of air and oil consumption
was possible. However, it was found that as the air
supply rate is reduced, oil begins to collect in

HSE(HSF) typeSF type

HSL(HSFL) typeFig. 5 Bearing design[Test Conditions]
Outer ring temperature ˚C

80[HSFL type]
5S-2LA-HSFL020 (φ100×φ150×24)

HSFL typeSF type
HSF type70[SF type]
5S-SF10XX (φ100×φ150×24)[HSF type]
5S-SLA-HSF020 (φ100×φ150×24)

60

50

0

10

20

30

40

50

Rotating speed :21000 min-1
Preload amount :2.5 kN (Constant preload)
Oil supply rate :0.03mL/5min
Jacket cooling :Provided

Air supply rate NL/min

Fig. 6 Air supply rate on outer ring temperature-22-

Precision Bearings "ULTAGE" Series for Machine Tools

a fixed position preload (mounted preload: 0N). With
the HSE type, the outer ring temperature becomes
unstable when the rotating speed increases over than
16,000 min-1, but the HSL type can achieve stable
operation at rotating speeds of up to 19000 min-1 (dmnvalue: 2,400,000), at 10NL/min (air supply rate) and at
0.03mL/10min (oil supply rate).

type, maximum reduction of oil consumption to 1/10 ispossible, since an oil supply interval of 2 min isnormally recommended for dmn value: 2,650,000.
Fig. 8 shows noise level change at different rotating
speeds of up to 20,000 min-1. It is demonstrated thatnoise level can be reduced by approximately 10dBA.
Fig. 9 shows operation results of the HSL type (5S2LA-HSL020: φ100×φ150×24) when operated with

[Test Conditions]
Outer ring temperature ˚C

80[HSFL type]
5S-2LA-HSFL020 (φ100×φ150×24)

HSFL typeSF type
HSF type70[SF type]
5S-SF10XX (φ100×φ150×24)

Unstable[HSF type]
5S-2LA-HSF020 (φ100×φ150×24)

60

50

0

5

10

15

20

Rotating speed :21,000 min-1
Preload amount :2.5 kN (Constant preload)
Oil supply rate :0.03mL/shot
Air supply rate :HSFL type: 12.5NL/min
SF type:25NL/min
HSF type:30NL/minJacket cooling :Provided

25

Oil supply rate min

Fig. 7 Effect of oil supply rate on outer ring temperature[Test Conditions]120

Noise level dBA[HSFL type]
5S-2LA-HSFL020 (φ100×φ150×24)

HSFL type
HSF type

110[HSF type]
5S-2LA-HSF020 (φ100×φ150×24)

10090Preload amount :2.5 kN (Constant preload)
Oil supply rate :0.03mL/5min
Air supply rate :30 NL/min
Jacket cooling :Provided

80
70
60

0

5000

10000

15000

20000

Rotating speed min-1

Fig. 8 Noise level[Test Conditions]
Outer ring temperature ˚C

65

55[HSL type]
5S-2LA-HSL020 (φ100×φ150×24)

HSL type
HSE type[HSE type]
5S-2LA-HSE020 (φ100×φ150×24)

Unstable

Preload amount :0N after assembled(Fixed position preloading)
Oil supply rate :0.03mL/10min
Air supply rate :10 NL/min
Jacket cooling :Provided

45

35

25
8000

10000

12000

14000

16000

18000

20000

Rotating speed min-1

Fig. 9 Outer ring temperature rise-23-

NTN TECHNICAL REVIEW No.71（2004）

bearing by supplying air and oil to the tapered portionof the inner ring, not directly to the rolling element.
This enables reduction of air and oil supply rates aswell as noise level.
Fig. 11 shows the bearing design of standard
bearing (N10HS type) and eco-conscious bearing(N10HSL type).
Fig. 12 shows the test results of outer ringtemperature rise for the N10HS and N10HSL types.
The N10HS type shows an acute temperature riseof outer ring temperature at 20,000 min-1 underconditions of 40NL/min (air supply rate), 0.02mL/5min(oil supply rate) and 0μm (clearance after mounted).
With the N10HSL type (N1014HSL: φ70×φ110×20), operation is possible under conditions of
20NL/min (air supply rate) and 0.02mL/10min (oilsupply rate), which are 1/2 of the conventional figures.
High-speed operation at 26,000 min-1 (dmn value:
2,300,000) is possible. From comparison of the outer

2. 1. 4 Eco-conscious Air-Oil LubricatedCylindrical Roller Bearing (N10HSL Type)
Conventionally, the allowed dmn value of air-oil
lubricated cylindrical roller bearings is approximately
1,500,000, therefore, angular contact ball bearings areused for the rear side of spindles that exceed dmn
1,500,000. In this case, a slide mechanism equippedwith a ball bushing is used to compensate spindle
extension for complicating the spindle structure (Fig.
10). To improve this structure , NTN developed the"Eco-conscious Air-Oil Lubricated Cylindrical Roller
Bearing (N10HSL Type)".
This bearing enables high-speed operation by
optimizing the internal structure and using a special
resin cage which is lighter than brass cage.
The lubrication system, like the eco-conscious
angular contact ball bearing, is constructed in such a
way that air cutting noise caused by the rolling
elements is reduced and oil is supplied inside of theFig. 10 ULTAGE main spindle rear structure is less complex than current rear structureStandard bearing (N10HS type)Eco-conscious bearing (N10HSL type)Fig. 11 Bearing design-24-Precision Bearings "ULTAGE" Series for Machine Tools

ring temperature between N10HS and N10HSL, it isobvious that the N10HSL type has a lower
temperature rise, which is because of the lowtemperature rise effect by smaller rolling elements. A
rise of outer ring temperature is shown at low speeds
for both types. This caused by the excessive supply ofoil which is suitable for the high-speed operation .
Fig. 13 shows the measurement results of noiselevel. The eco-conscious type shows a lower noiselevel by approximately 6dBA than the standard type at
15,000 min-1.

Outer ring temperature rise ˚C[Test Conditions]
16[N10HSL type]N1014HSL (φ70×φ110×20)

N10HSL type
N10HS type14
12

Clearance
Oil supply rate
Air supply rate
Jacket cooling

10
8
6[N10HS type]N1014HS (φ70×φ110×20)

4Clearance
Oil supply rate
Air supply rate
Jacket cooling20:0 μm after mounted:0.02mL/10min:20 NL/min:Provided

0

5000

10000

15000

20000

25000

30000

Rotating speed min-1:0μm after mounted:0.02mL/5min:40 NL/min:Provided

Fig. 12 Outer ring temperature rise

[Test Conditions]120

Noise level dBA[N10HSL type]N1014HSL (φ70×φ110×20)
Clearance:0μm after mounted
Oil supply rate :0.02mL/10min
Air supply rate :20 NL/min
Jacket cooling :Provided

N10HSL type
N10HS type

110
100
90
80
70
60

0

5000

10000

15000

20000

25000

Rotating speed min-1[N10HS type]N1014HS (φ70×φ110×20)
Clearance:0μm after mounted
Oil supply rate :0.02mL/5min
Air supply rate :40 NL/min
Jacket cooling :Provided

Fig. 13 Noise level-25-

NTN TECHNICAL REVIEW No.71（2004）

of improving resistance to fretting by applying surface
heat treatment and pre-lubricating with special grease(the oscillation test results show the improvement in
fretting resistance by 10 times). The results are shownin Fig. 15.
Unlike the BST type standard bearing (opendesign), the 2A-BST LXL type that has low-torque,
light-contact seals, which not only prevents entry ofcontaminants but also facilitates assembly work since
the greasing process is not required. Furthermore,
confirmation of bearing orientation during assembly isalso made easier by changing the seal colors betweensides, like the sealed angular contact ball bearing
developed for main spindles.

2. 2 Ball Screw Support Bearing(2A-BST LXL Type)To improve efficiency of machine tools, the speed of
the feeder also needs to be improved in addition to themain spindle. Ball screw support bearings used forfeeders sometimes cause fretting on the raceway due
to the oscillation generated by slight feed movement
and vibration generated during cutting operation. In
addition, deterioration of grease due to entry of cutting
oil and shortened life due to entry of contaminants(e.g. ball screw grease, cutting chips) are alsopresent.
The ball screw support thrust angular contact ballbearing (2A-BST LXL type) (Fig. 14) has beendeveloped to solve these problems, with the intention

Black seal
Orange seal

Fig. 14 Thrust angular contact ball bearing for ball screw support

1.2

Load

Maximum contact stress :1.7GPa
Oscillation angle:12 deg
Oscillation cycle:30 HzDuration:8 hAbrasion ratio

1.0
0.8
0.6
0.4
0.2

Rotary oscillation

0

BST type

Fig. 15 Oscillation test results-26-

2A-BST LXL
type

Precision Bearings "ULTAGE" Series for Machine Tools

3. Conclusion
Machine tools in the future are expected to feature
eco-consciousness in addition to conventional
functional requirements such as high speed, highefficiency and high precision. Eco-conscious
technologies, including reduction of air/oil
consumption, oil mist and noise level with powersaving will play more important roles.
For machine tool bearings, including those forspindles and feeders, eco-consciousness will be a key
issue in addition to the quest for higher dmn value andlower NRRO.
A wide range of requirements is expected from
customers. However, NTN will continue to work on
improvement and development of ultimate precision
bearings keeping its motto "be friendly to the earth
and its people" in mind.

References
1) H.Tako , THE TRIBOLOGY NO.164(2001),24 (inJapanese)2) K.Ueda , THE TRIBOLOGY NO.188(2003),19 (inJapanese)
3) H.Takiuchi , THE TRIBOLOGY NO.174(2002),58 (inJapanese)
4) K.Fujii , M Mori , JSPE Semestrical meeting(autumn)(2001.10) (in Japanese)5) K.Fujii , JSME Division of Manufacturing and Machine
Tools
The 3rd International Conference (2001.11) (inJapanese)

Photo of the author

Futoshi KOSUGIIndustrial Engineering Department
Industrial Sales Headquarters-27-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Bearings for High Speed CT Scanner

Yosuke OYA*The Industry of Medical Instruments is steadily growing because of an aging society.
In particular, the CT(Computed Tomography) scanner, which can be an effective tool for medical examination andinspection of patients, has attracted a attention as important imaging equipment.
To ensure smooth rotation of the gantry part of the CT scanner, a super slim large size angular contact ball bearing is
used.
This article introduces development of the gantry bearing for higher speed rotation with a lower noise level.

1. Introduction
The medical instrument industry is steadily growing
due to an aging society. In particular, the CT(Computed Tomography) scanner has attracted
attention as imaging equipment to enable efficient
medical examination and inspection of patients. The
CT scanner (Photo 1) has an inspection section(gantry) that houses an X-ray tube and detector as

part of a rotating section to shoot images. A bearing(inner diameter: approx. 1m) is used to support thisrotating section. This article introduces thedevelopment of rotation support bearings (gantry
bearings) that can support improvement of CT
scanner performance.

Gantry part

Bed

Photo 1 CT scanner*Industrial Engineering Department Industrial Sales Headquarters-28-

Bearings for High Speed CT Scannercontact ultra-thin ball bearings are used for low- to
medium-speed CT scanners because of theircompactness. For medium- to high-speed CT
scanners, duplex double-row angular contact ballbearings are used to reduce heat build-up inside thebearing and keep rigidity. In particular, duplex doublerow angular contact ball bearings, which are used forhigh-speed CT scanners, have the following features.

2. Required Functions for Gantry
Bearings
Required functions for gantry bearings are
explained below.

(1) Higher-speed operation
Improved speed of gantry bearings shortens
imaging time, resulting in reduced burden on
patients. In other words, the time during which thepatient has to stop breathing can be reduced andthis reduces the burden placed on patients,
especially infants and the elderly. Moreover,reduction of exposure to X-rays is also an
important purpose of speed improvement.
Furthermore, higher rotation of the bearing
improves imaging speed, enabling high precision
imaging of internal organs (e.g. heart), which weredifficult to shoot accurately. Furthermore, thenumber of patients for whom diagnoses can bemade per day increases, thereby resulting in ahigher operation availability of the CT scanner.

(1) Preloading specification
If there is clearance inside the bearing, non-loadstate will exist while the rolling elements make oneturn inside the bearing during rotation. In this nonloaded area, a striking sound will be generated as
the rolling elements collide with the ring by theirown weight. Since gantry bearings are required tobe silent, the bearing we have developed is thepreloaded type that prevents generation of suchstriking noise.
Fig. 2 shows the state in which clearance isexistent inside the bearing when moment load isexerted on the bearing, and the state in which an
appropriate preload is exerted on the entire
bearing. The horizontal axis of the graphs indicatesthe positions of rolling elements (e.g. 110 balls)
located along the entire inner circumference of the
bearing, and the vertical axis indicates the contact
stress on the rolling elements and rings. Setting anappropriate preload eliminates non-load state on
the rolling elements across the entire inner
circumference of the bearing. In addition, exertionof appropriate preload keeps required bearing
rigidity and improves imaging accuracy.

(2) Low noise level
Reducing noise level of gantry bearings can not
only ease patients' anxiety during image shooting,but also prevent atrophy of organs.

3. Construction
Fig. 1 shows the construction of a gantry bearing.
Since the gantry bearing rotates while supporting thebase to which imaging devices (e.g. X-ray tube anddetector) are attached, moment load will act on the
bearing. In general, single-row type, four-point angular

Single-row typeDouble-row typeFig. 1 Schematic of CT scanner bearing-29-

NTN TECHNICAL REVIEW No.71（2004）

2000

2000
Inner ring raceway (Bearing !)Contact stress (MPa)Contact stress (MPa)

Inner ring raceway (Bearing!)
Outer ring raceway (Bearing @)1500
Inner ring raceway (Bearing!)
Outer ring raceway (Bearing @)

1000

500

Outer ring raceway (Bearing @)1500

Inner ring raceway (Bearing !)Outer ring raceway (Bearing @)

1000

500

0

0
0

55

0

110

55

110

Position of rolling element

Position of rolling element

When clearance is existent inside bearing

When appropriate preload is exerted on bearing

Fig. 2 Relationship of contact stress between each ball position and inner/outer ring(2) Development of low-noise cage (Table 1, Fig. 3)

the cage and inner- outer-ring during rotation.
Furthermore, the rigidity of the cage has been
improved by employing a PPS resin cage insteadof the conventional PA resin cage and improving
GF to 30%. The elimination of the opening was theresult realized by consideration given to thecharacteristic of the PPS resin so that it does notcause dimensional change through water
absorption.

An especially important factor in noise reduction is
optimization of cage design. In particular, large,
thin cages, like those used for gantry bearings, arethe link type and made of resin.
Conventionally, PA material was used as theresin, but it caused dimensional changes due toabsorption of water, resulting in interference withthe outer/inner rings and generation of stress on
the cage itself. Therefore, a free area (opening)
was provided, instead of connecting the entire
circumference of the cage.
However, since this opening was not constrained
inside the bearing, collision sounds (striking noise)
between the inner and outer rings during rotation
was generated, resulting in being one of thecauses of unpleasant jarring sound on the ear.
The cage we developed has no such opening
and has its entire circumference connected toeliminate such striking noise, and employs amechanism that guides the cage by means ofrolling elements to prevent interference between

Table 1 Comparison of cage type
Conventional cage
Developed cage
Opening
Not provided
ProvidedMaterial
PPS+GF30%
PA66+GF10%
Pocket form Cylindrical and square (alternately)Special spherical
Guide typeRolling element guide
−

Sound

Aural judgment: × (NG) Aural judgment: ○ (OK)○ No striking sound from× Striking sound from
openings
openings
× Striking sound between ○ No striking soundbetween cage and rings
cage and rings(○ No dropping sound of(× Dropping sound ofrolling elements)
rolling elements)｛

Opening (One on circumference)
Conventional cage

Developed cage

Fig. 3 Cage forms-30-

Bearings for High Speed CT Scanner

4. Confirmation Test Results for
Required Functions(Test Results)

The results of each test conducted on the required
functions of gantry bearings are explained below.

Fig. 4 shows the relationship between bearing
rotating speed and outer ring temperature rise.
With current high-speed CT scanners, the bearing
rotating speed varies from 120 to 180 min-1,
depending on the bore diameter of the gantry. It
was confirmed by this test that no abnormal outerring temperature rise occurred even at 300 min-1(dmn=270,000) and operation was stable at highspeeds twice those of current speed.

(1) Higher-speed operation(Test Contents)
A double-row angular contact ball bearing(approx. ID800 ×OD1000×Width 60 mm) was
attached to the back plate of the test equipment
shown in Photo 2. A weight was placed
approximately 150 mm away from the center of the
bearing, and test was conducted with radial load of
8000N. Evaluation was made at three bearing tilt
angles : 0˚ (upright) and ±30˚ (tilted to front andrear), which are normally used for CT scanners.

Photo 2 Test equipment

Bearing outer ring temperature (˚C)

10Upright
Tilted by 30° towards front

8

Tilted by 30° towards rear
6

4

2

0
0

50

100

150

200

250

Bearing rotating speed (min-1)

Fig. 4 High speed rotation test results

-31-

300

350

NTN TECHNICAL REVIEW No.71（2004）

level is 72dBA at 60 min-1. Furthermore,
unpleasant sounds on the ears, which are typicallyheard in the case of cages that operate insynchronization with the bearing, were not heard.
Fig. 6 shows the waveform of noise obtained
when the test was conducted at 120 min-1. In thecase of the bearing equipped with the conventional
cage, the noise waveform shows clear peaks
synchronized with rotation, and at that time peaksin the range of 1000 to 1500 Hz in particular were
heard as noise. On the contrary, with the bearing
equipped with the newly developed cage, no peaks
synchronizing with rotation were observed, and it
was confirmed that noise level was low on thewhole.

(2) Low noise level(Test Contents)
A double-row angular contact ball bearing(approx. ID1000×OD1200×Width 80 mm) was
attached to the back plate of the noise level test
equipment shown in Photo 3. A microphone wasplaced approximately 1m away from the center ofthe bearing, and test was conducted.

(Test Results)
As shown in Fig. 5, the noise level increases
proportionally to the bearing rotating speed.
However, with the bearing equipped with thedeveloped cage, the noise level is 3dBA lower,
even at 180 min-1, than that of the bearing
equipped with the conventional cage, whose noise

Noise level (dBA)80
78

Conventional cage

76

Developed cage

74
72
70
68
66
64
62
60

60

120

Rotating speed (min-1)

Photo 3 Test equipment

Fig. 5 Noise level test rsesults

Developed cage

Conventional cage
-1)

Fig. 6 Wave form of noise (120min

-32-

180

Bearings for High Speed CT Scanner

5. Conclusion
In response to improved speed of gantry bearings,
we were able to develop bearings that not only enablehigh-speed operation but also provide low noise level,by setting an appropriate preload and employing a
cage of appropriate material and form in double-row
angular contact ball bearings. Development of thisbearing will surely help in improving the speed andreducing the noise level of CT scanners. From now
on, we will also continue to develop bearings that willcontribute to the entire medical instrument industry.

Photo of the author

Yosuke OYA
Industrial Engineering Department
Industrial Sales Headquarters-33-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Integrated Sensor Bearing Unit for Axleboxes

Masanori UENO*

Constant monitoring of axlebox bearings , which are some of the most important components of rail vehicles,
contributes to the reliability and safety of railways. In addition, railway companies today desire to reduce components
and to simplify maintenance procedure for reduction of maintenance cost.
To meet this demand , we have developed a sealed double row tapered roller bearing with an integrated sensor. Thissensor is incorporated into plastic housing which is attached to the oil seal case on the shaft end side. This sensor can
detect temperature , rotating speed and rotating direction.
This report introduces laboratory test results for performance of the temperature sensor and impact- and vibrationresistance of the speed sensor.

1. Introduction
Railroad vehicle axle bearings are very important.
Damage to them will hinder proper operation of therailroad vehicle and may even result in serious
accidents. Due to recent extensions of maintenanceintervals and increases in vehicle speed, a higher
level of reliability is required of these bearings.
Railroad axle bearing operating parameters are
currently monitored through constant observation oftemperature using sensors attached to the axle box orby periodic inspection of the thermo-label affixed tothe axle box/axle end. In addition, detection of axle
rotating speed and associated wheel sliding during
braking has been conducted using a gear attached tothe axle end and a speed sensor attached to the axle
box. However, this method requires many large
components requiring maintenance adjustments of thegap between the sensor and gear.

Since the demands for reduction of maintenancecosts are high for railroad vehicles, simplification ofaxle bearing and reduction in the number ofcomponents are desired. The development of an
integrated sensor-bearing unit that enables detectionof both temperature and rotating speed of bearings isalso desired. NTN has developed an axle bearing witha sensor function, consisting of a sealed double row
tapered roller bearing that is widely used for railroad
vehicles. This paper introduces this new axle bearing
together with various evaluation tests that were
performed.

*Industrial Sales Headquarters Industrial Engineering Department-34-

Integrated Sensor Bearing Unit for Axleboxes

The magnetic ring consists of N and S poles
magnetized alternately in the circumferential direction.
This causes the magnetic field to pass through theHall ICs according to the axle rotation. The speed
sensors convert the change in the magnetic field tovoltage pulses. Two Hall ICs are placed so that thephase difference of the output pulses is 90 degrees.
By detecting the frequency and phase difference ofthe pulses, the rotating speed and direction of the axle
can be found.

2. Structure of Sensor Axle Bearing Unit
The bearing unit developed by NTN consists of asealed, double row, tapered roller bearing with an
inner diameter of 120mm and outer diameter of
220mm and a sensor unit. The sensor unit is attachedto the oil seal case on the axle end side, and designed
to detect temperature and rotating speed/direction of
the inner ring (see Fig. 1).
To maintain compatibility with conventional axle
bearings, the external dimensions of the unit are kept
the same as those of conventional axle bearings.
Specially shaped oil seal lips have been designed tothat offer sufficient installation space of the sensor
unit. The sensor unit is bolted to the seat on the oilseal case. Since axle bearings for railroad vehicles
are exposed to operational impact and vibration fromrail tracks, specially designed jam washers are used
to prevent loosening of bolts.
In addition, the magnetic ring used for detection ofrotating speed is fit onto the rib of the inner ring on theaxle end side as shown in Fig. 1. The magnetic ring
can also be attached to the seal wear ring dependingon the structure of the bearing.

Temperature sensor

3. Specifications of Sensor Unit
Photo 1 shows an external view of the sensor unit,
while Table 1 shows the main specifications.
The sensor unit comes in a resin housing containing
a temperature thermistor (thermally sensitive resistors)
to measure the bearing temperature. Two magnetic
sensor elements (Hall IC) are used to detect the axle
rotating speed.

Speed sensor

Photo 1 Sensor unit

Sensor unit

Magnetic ring

Fig. 1 Axlebox bearing unit with integrated sensor-35-

NTN TECHNICAL REVIEW No.71（2004）

Table 1 Specification of sensor unit
Operating temperature range
Impact resistanceVibration resistanceFrequency response of speed sensor
Adjacent pitch error
Resolution of
speed sensor
Duty ratio(Fig. 2)Phase difference between A & B signals

In this test, four sensor units were attached to the oilseal case, one at every 90 degrees in thecircumferential direction, to check the difference in the
output temperatures due to different sensor unit positions.

-40 to 125˚C
100 G
35 G
0 to 8kHz
5% or less
50±15%
90±45 degrees(Test Results)
Fig. 4 shows the outputs from the temperature
sensors and the results measured by the thermocouplers.
¡The outputs from the temperature sensors areapproximately 10˚C lower at maximum than thetemperature measured on the outside surface of theouter ring by the thermocouple. However, they canmonitor temperature change on the outer ring,while
the temperature sensor is available for actual use.
¡Comparison of the sensor outputs with eachposition indicates that the output from the top
sensor (in the bearing load range) is approximately
5˚C higher at maximum than the others. There isalmost no difference among the other outputs.

Phase A
TAB

Phase B

TP
Tn

Tn+1

1Adjacent pitch error (%)＝｜（Tn−Tn+1）｜Tn×100
2Duty ratio (%)＝TP/ Tn×100

100

Temperature output (˚C)

3Phase difference between A & B signals (degree)＝TAB/Tn×360

Fig. 2 Resolution of speed sensors

4. Evaluation Test
4. 1. Bearing Temperature Detection Test

90
80
70
60

Outside surface of O.R. (thermo-coupler)
End surface of O.R. (thermo-coupler)Sensor (top)Sensor (bottom)
Sensor (right)
Sensor (left)

50
40
30

The axle bearing test rig shown in Fig. 3 was used
to test the axle bearing sensor unit. The output
temperature of the temperature sensors wascompared with the bearing temperature measured bya thermocouple attached to the bearing.

20
0

100

200

300

400

500

Time (min)Fig. 4 Relationship between bearing temp. and output oftemp. sensor

Load cell
Radial load cylinder
Test bearing
Support bearing
Axial load cylinder

Load cellFig. 3 Test rig for axle bearings-36-

Integrated Sensor Bearing Unit for Axleboxes

4. 2. Vibration Test
The sensor units were exposed to continuous
vibration. The test confirmed the resolution of the
speed sensors shown in Table 1 was maintained.

Adjacent pitch error (%)

6(Test Conditions)¡Vibration acceleration：±35G¡Vibration direction ：Vertical, crosswise, lengthwise¡Frequency：60Hz¡Vibration cycle：107 cycles for each direction¡Specimen：3 specimens

5

Upper specification limit of resolution

4
3

Blue: Phase A
Red: Phase B2
10(Test Results)

VerticalFigs. 5 to 7 show the measured adjacent pitch
error, duty ratio and phase difference between A & B
signals, respectively. The designed resolution of the
speed sensors was verified even at the end of thetest, demonstrating that the sensors have sufficientvibration resistance required for railroad applications.

Crosswise

Lengthwise

Fig. 5 Adjacent pitch error

100

Blue: Phase A
Red: Phase B

Duty ratio (%)

80

Upper specification limit of resolution
60

40

Lower specification limit of resolution
20

0

Vertical

Crosswise

Lengthwise

Phase difference between A & B signals (degree)

Fig. 6 Duty ratio

150
Upper specification limit of resolution
120

90

60

Lower specification limit of resolution
30
Vertical

Crosswise

Lengthwise

Fig. 7 Phase difference between A & B signals-37-

NTN TECHNICAL REVIEW No.71（2004）

4. 3. Impact Test
The sensor units were exposed to impact. The test
confirmed the resolution of the speed sensors shownin Table 1 was maintained.

6
Upper specification limit of resolution
Adjacent pitch error (%)

5(Test Conditions)¡Impact：100G¡Impact direction ：Vertical, crosswise, lengthwise¡Vibration cycle ：4000 cycles for each direction¡Specimen：2 specimens

4
3
Blue: Phase A
Red: Phase B
2
1(Test Results)

0

Figs. 8 to 10 show the measured adjacent pitch
error, duty ratio and phase difference between A & B
signals, respectively. The designed resolution of the
speed sensors was verified even at the end of thetest, demonstrating that the sensors have sufficient
impact resistance required for railroad applications.

Vertical

Crosswise

Lengthwise

Fig. 8 Adjacent pitch error

100

Blue: Phase A
Red: Phase B

80
Duty ratio (%)

Upper specification limit of resolution
60

40
Lower specification limit of resolution
20

0
Vertical

Crosswise

Lengthwise

Phase difference between A & B signals (degree)

Fig. 9 Duty ratio

150
Upper specification limit of resolution
120

90

60
Lower specification limit of resolution
30
Vertical

Crosswise

Lengthwise

Fig. 10 Phase difference between A & B signals-38-

Integrated Sensor Bearing Unit for Axleboxes

5- Conclusion
This overview of the integrated sensor-bearing unit
designed for axle boxes with associated test results
has been presented in response to an expected
increase in demand for improvement in safety andreliability of railroad vehicles. Additionally, railway
companies are likely to seek reduction of vehicle
maintenance costs. Reliability of railroad axle bearings
is critical. NTN is confident the integrated sensorbearing unit introduced in this paper will be aneffective means of addressing this attribute while
helping to reduce the size and maintenance cost of
the bearings.

Photo of the author

Masanori UENO
Industrial Engineering Department
Industrial Sales Headquarters-39-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Bearings for Wind Turbine

Souichi YAGI*In 2002, worldwide electricity production was about 31,000MW. This is a 27％ increase over the previous year.
In the past few years, the wind turbine generating system, which emits no carbon dioxide, has gained widespread
acceptance as the cleanest and most environmentally friendly energy. The technical trend for wind turbines is toincrease reliability and efficiency while reducing the cost of operation. The bearings, which are one of the mostimportant components for wind turbines, require designs that optimize reliability and economic efficiency whileconsidering the characteristics of this applications.
This report introduces special characteristics for wind turbine bearings and a method to optimize wind turbine
bearing design.

1. Introduction

2. Structure of Wind Turbines andBearings

Worldwide electricity production reached about
31,000MW by the end of 2002. This is a 27％ increaseover the previous year. In the past few years wind
power generation, which emits no carbon dioxide, hasgained widespread acceptance as the cleanest andmost environmentally friendly form of energy.
Technical issues remaining in the development ofwind power generation include increasing reliability ofthe system while reducing the cost of operation.
Bearings are one of the most important components ofwind turbines and require designs that optimize
reliability and economic efficiency while taking into
account the characteristics of the applications. This
report introduces a method to optimize wind turbine
bearing design and the features of bearings
developed for wind turbines.

Fig. 1 shows the nacelle of 1 to 2MW wind turbines.
Bearings are used in various places of the nacelle:rotor shaft, gearbox (step-up gear), generator, yaw
gearbox (reduction), yaw slewing table, blade pitch
revolving seat and hydraulic pump.

3. Bearing Operating Conditions
The rotor shaft bearing supports the blades androtor and transmits torque to the gearbox. The bearing
loads and rotating speeds vary considerably due toconstantly changing winds.
At wind speeds below the cut-in wind speed (i.e. theminimum wind speed required for power generation),
the rotor shaft will idle resulting in low-speed, low-load
operation. At wind speeds above the cut-in speed, the*Industrial Sales Headquarters-40-

Bearings for Wind Turbine

Rotor shaft

Gearbox (step-up gear)

Blade

Generator

Yaw gearbox (reduction gear)

Fig. 1 NacelleRotor shaft bearings repeat start, acceleration,
deceleration and stop operations irregularly as they
are exposed to fluctuation of load. Therefore, theoptimal specifications for various parameters,
including bearing type, clearance, number of bearing
rollers, crowning and cage must be examined for eachcondition (minimum load, average load, maximumload).

rotating speed increases above the rated speed,
resulting in average loads. In the case of wind gusts,
the blades and rotor will exert large loads on the rotor
shaft bearing. Fig. 2 shows the load and moment
exerted on the rotor shaft bearing.
Such changes in the load, moment and rotating
speed also affect the gearbox bearing. One of thefeatures of wind turbine bearings is that they operate
in a wide range of loads from light to heavy load(when exposed to gusts). Fig. 3 shows examples ofmeasured rotor load and moment over one minutewhen a 700-kW wind turbine is operated at windspeeds of 23 to 24m/s.

140

FXN

100

（kN）

60

MZN

250

FYN

0
-250

FZN

130

FZN

MXN
100

MYN

FXN

ZN

0

FYN
YN

XN

10

20

30

40

50

280

60（秒）

MXN

（kNm）

220
160
300

MYN

0
-300
200

F ：Load
M：Moment

MZN

0
-200
0

Fig. 2 Rotor load schematic

10

20

30

40

50

Fig. 3 Measurements of rotor load and moment
-41-

60（秒）

NTN TECHNICAL REVIEW No.71（2004）

When designing the bearing, first calculate thebearing life for the maximum required strength of the
housing and the average deformation of the housing.
Then design a slim housing and choose a shaft
bearing that meets the required calculated life.
With deformation of the housing and outer ring
raceway taken into account, calculate the load oneach rolling element to obtain the life of the rotating
and stationary rings.

4. Rotor Shaft Bearings
Table 1 shows the structures of the shafts that use
a gearbox to increase blade speed to the rated speedof the induction generator. Bearings suitable for eachrotor shaft type are also shown. Table 2 shows the
structures of the shaft of synchronous generators not
equipped with a gearbox.

Table 1 Wind turbine rotor shaft bearing assembly(with gearbox)
Structure

Blade-side Generator-side
bearingbearing¡Two bearings are
used.

SRB
CRB
DTRB

SRB

CRB

¡The generator-side
bearing is also usedas the gearbox's
input bearing.

CRB¡The generator-side
bearing is also usedas the gearbox's
input bearing.
¡The load on theblade-side bearing
is supported by thenacelle.

TRRB
DTRB

SRB : Spherical roller bearing
DTRB : Double-row tapered roller bearing―

QR , Qs：Average load on rotating andstationary rings
Z：Number of rolling elements
wi, we：Constantp
L R＝（Cn／QR）：Life of rotating ringp
LS＝（Cn／Qs）：Life of stationary ring-e-1/e
L＝（LR ＋LS-e）：Life of bearing
Cn ：Dynamic rated load on contact point
p ：For ball bearing 3
For roller bearing 10/3
e ：For ball bearing 10/9
For roller bearing 9/8¡The gearbox is
supported on therotor shaft.

Spherical roller and tapered roller bearings aremainly used. However, spherical roller bearings, thatfeature low misalignment rates, are widely used.
Normally, misalignment of ±0.5˚ needs to be taken
into account.
With B type spherical roller bearings, the rollers are
guided by the inner ring's center rib. This enables
operation with stable torque, low skew and low heatgeneration in a wide load range from the minimum tothe maximum load. Fig. 4 shows the structure of
NTN's spherical roller bearings (both B and C types).

¡No rotor bearing isused and the rotor
load is borne by the
gearbox bearing.

CRB : Cylindrical roller bearing
TRRB : Triple-row cylindrical roller bearing

Table 2 Wind turbine rotor shaft bearing assembly(without gearbox)
Structure

Blade-side Generator-side
bearingbearing

TRRB
DTRB

CRB

SRB
DTRB

CRB
CRB

we
1 Z
Σ（QSJ）
Z J＝1

1/ we

QS＝

Features

SRB
SRB
SRB

SRB

wi
1 Z
（QRJ）
Σ
Z J＝1

1/ wi

QR＝

Features¡Direct drive¡Outer ring rotation¡The load on theblade-side bearing
is supported by thenacelle.

B type¡Inner ring rotation

C typeFig. 4 Spherical roller bearing, B type and C type-42-

Bearings for Wind Turbine

Fig. 5 shows an example of the deformation of arotor shaft bearing outer ring and housing. The
bearing and housing are optimized by calculating thebearing life, including housing deformation andbearing clearance and also confirming the housing
strength based on a stress analysis. In this example, abearing for 1.5-MW wind turbines was used, and the
maximum deformation in the axial direction was
0.07mm. The difference in the bearing life calculatedwith this deformation and bearing clearance taken into
account and the one calculated with the housing andouter ring considered to be non-deformed was within
5%, which is not problematic in actual use. If thisdifference is excessively large and the life is short, thedesign of the housing needs to be changed to improverigidity.

Since rotor shaft bearings are exposed to vibrationof blades and gearbox, fretting corrosion may occur.
Therefore, selection of appropriate bearings andgrease, as well as optimization of clearance and fitting
are important factors.
Photo 1 shows an external view of a bearing at theend of vibration test.

(1) Fretting on inner ring bore

Fig. 5 Bearing outer ring and Pillow Block deformation(2) Fretting on inner ring raceway

5. Planet Bearings for Gearbox《Vibration test》
Bearing：24126CL1
Vibration acceleration：9G
Vibration cycles：10 million cycles
Maximum surface pressure：1080MPa

A gearbox consists of an input shaft, planet gear,
low-speed shaft, intermediate shaft and high-speed
shaft. Fig. 6 shows the structure of a gearbox, andPhoto 2 shows an external view of a full complementcylindrical roller bearing, that is used for input and lowspeed shafts.
Fig. 7 shows an example of planet gear
mechanism.

Photo 1 Vibration test results-43-

NTN TECHNICAL REVIEW No.71（2004）

Spherical roller bearings or full complementcylindrical roller bearings are used as planet bearings.
An analysis model of the planet bearing is used to
calculate the life of bearing with deformation of theouter ring taken into account. When calculating, makesure that the load on the bearing is considered to be
the one exerted on rolling elements and the gear
engagement point between the ring gear and sun gearis fixed.
Fig. 8 shows an example of deflection analysis
results of a planet bearing used for 1.5-MW gearbox.
The planet bearing used for analysis consists of two
double-row cylindrical roller bearings with four rows ofrolling elements, and the maximum deflection was
0.21mm.

High-speed
shaft

Input shaft
Intermediate
shaft

Low-speed
shaft
CarrierPlanet bearing

Planet gear

Support bearing

Fig. 6 Gearbox for wind turbine

Fig. 8 Deflection of planet bearing

The life, calculated from the analysis results withelastic deformation of the rings taken into account,varies by a maximum of 58% among rows. The
carrier-side rows have higher load ratio than theothers, resulting in shorter calculated life. Fig. 9
illustrates the arrangement of bearings, and Table 3
shows the calculated results.

Photo 2 NTN Full complement cylindrical roller bearing

Planet gear
1

Carrier2

3

4

Generator side

Carrier side

Ring gear
Fig. 9 Arrangement of bearings
Sun gear(Low-speed shaft)

Table 3 Life ratio of each row for planet bearing
Row No.

Fig. 7 Planetary gear model-44-

Life ratio

1

85

2

143

3

133

4

100

Inner diameter φ220mm
Double-row cylindrical roller bearingBearings for Wind Turbine

The maximum contact surface pressure (Pmax) for
gearbox bearings is calculated based on point contact
for spherical roller bearings, and line contact forcylindrical roller bearings. In most cases misalignmentis taken into account when specifying the maximum
contact surface pressure.

implemented to prevent a situation in which the planet
bearing is exposed to insufficient lubrication when it
begins to move.
The problems concerning the planet bearings canbe summarized into the following three points.
¡Influences by plastic deformation of gears andbearing¡Influences by misalignment of planet bearing
caused by twisting of carrier¡Influences by dry start up (insufficient lubricatingoil)
The specifications for the optimal bearing must be
take into account these properties. NTN has beenworking on the design of ribs that have high load
capacity and that provide sufficient resistance against
sliding and scuffing under light loads, as well optimal
axial clearance. In addition, NTN has also beenworking to prolong the life of bearings by employing
special heat treatment.

Pmax＝Krc Km Pline (Cylindrical roller bearing)
Krc ：Crowning correction factor
Km ：Misalignment factor
Pline ：Maximum line contact surface pressure
In the case of spherical roller bearings, if the
maximum contact surface pressure exceeds the limit,
the bearing size needs to increase. This will increase
the calculated life and reduce the maximum contact
surface pressure. However, in the case of light loads
the rolling elements may not roll properly on the
raceway and will begin to slide. This sliding may
cause damage to the raceway. Because of this, NTN
recommends the minimum load be at least 4% of thebasic static load rating.
Concerning lubrication, some measures need to be

Furthermore, NTN has introduced the special test
machines shown in Figs. 10 and 11, to promote thedevelopment of next generation planet bearings.

1990

Test bearing

2150

1450

Fig. 10 NTN outer ring rotating test machine-45-

NTN TECHNICAL REVIEW No.71（2004）

1850

Test bearing

1400

1300

Fig. 11 NTN inner ring rotating test machine

6. Insulated Bearings for Generator

7. Bearings for Yaw Gearbox

To improve the reliability of bearings used ingenerators, it is necessary to prevent sparks (galvanic
corrosion) caused by electric current passing throughthe bearings. NTN has produced a new single-layer
bearing having sufficient insulation capability andreliability by adopting special ceramics and improving
the spray-coating method.
This bearing provides insulation resistance of 100MΩ
or higher and dielectric breakdown voltage of 2kV or
higher, meeting the insulation performance required
for wind turbines.

Since yaw gearboxes need to be small and capableof conveying large torque, the bearings to be used forthem must be compact and have high load capacity.
Because of this, angular contact ball bearings with thin
inner/outer rings and tapered roller bearings are often
used.
With angular contact ball bearings, which areexposed to large axial load, the resistance to axial
load has been improved by increasing the groove
depth on the inner and outer rings. Fig. 12 shows the
cross-sectional view of the standard bearing and thatof special design bearings for yaw gearboxes. The
special design bearing has a high axial load
resistance, approximately 9 times higher than thestandard bearing.

Photo. 3 shows an external view of this insulated
bearing.
For details, refer to "Insulated Bearing "MEGAOHM"
Series" in this book.

Standard bearing

Photo 3 Insulated Bearing

Special design bearings
for yaw gearbox

Fig. 12 Special design angular contact ball bearings-46-

Bearings for Wind Turbine

For tapered roller bearings, carbonitrided ETAbearings are used to provide longer life.
ETA bearings are long-life bearings with highthermal stability gained by optimizing distribution ofretained austenite and carbide present on the surface
through special heat treatment. They feature highresistance to contaminants contained in lubricating oil
and high peeling resistance. Table 4 shows the life
test conditions and Fig. 13 shows the results.
Such implementation has enabled NTN to provide
compact, highly reliable bearings.

8. Conclusion
Compared with Europe and North America, where
the use of wind turbines is widespread, Japan often
suffers from considerable atmospheric turbulence,
severe tropical storms and lightening (winter). Thus,
wind turbines that are reliable and suit Japan's climaticconditions are desired. NTN has been working on
improving the reliability of bearings designed for windturbines. By choosing optimal bearing specifications
that satisfy wind turbine manufacturers and users andsupplying high-quality products, NTN will contribute to
the development of an eco-friendly wind power
generation.

Table 4 Test conditions (30206, ETA30206)
Normallubricating oil

Contaminatedlubricating oil (Reference)

Radial load (kN)

17.64

Rotating speed (min-1)

2000

Lubricating oil

Turbine oil 56

References
1) Germanischer Lloyd
Regulations for the Certification of Wind Energy
Conversion Systems
2) B.Schlecht
Moderne Simulationstechniken zur dynamischen
Auslegung von Triebstraengen in Multi- MegawattWindenergieanlagen3) B.Schlecht et al"MULTIBODY-SYSTEM-SIMULATION OF DRIVE
TRAINS OF WIND TURBINES"
4) B.Niederstucke et al
LOAD DATA ANALYSIS FOR WIND TURBINEGEARBOXES

Turbine oil 56+
NTN's standard contaminants

Cumulative percent failed (%)

99

Standard bearing
標準軸受
ETA bearing
ETA軸受

80
50
20
10
5

1

100

101

102

103

Life (h)Fig. 13 Comparison of life of ETA tapered roller bearing
and standard bearing (with contamination)

Photo of the author

Souichi YAGI
Industrial Sales Headquarters-47-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Insulated bearing "MEGAOHM" series

Hideji ITO*NTN has developed a new ceramic-coated, insulated bearing that uses a single-layer, spray-coated ceramic.
Applications for this bearing include electrical wind turbine generators, general purpose motors, and traction motors
for trains.
NTN recently established the MEGAOHM series of insulated bearings. The MEGAOHM series, includes multi-layer,
spray-coated ceramic, insulated bearings, the new single-layer type, and the PPS resin-coated type.
The new single-layer, ceramic-coated, insulated bearing is introduced below.

1. Introduction
Traction motors for trains use bearings that are
insulated to prevent electric corrosion by coating theouter diameters with ceramic or injection-molded resin(glass fiber contained).
Conventionally, bearings were coated with ceramic
in multi layers (3-layer, 2-layer). The new bearing
developed by NTN (Fig. 1) is a low-cost, single-layer,ceramic-coated, insulated bearing and is designed foruse in wind turbine generators, general-purpose
motors and traction motors for trains. This paper
explains the bearing features and evaluation test
results.
In addition to the conventional, multi-layer, ceramiccoated, insulated bearings and special, resin-coated,insulated bearings, the bearings developed this time
are named "MEGAOHM Series", featuring highinsulation performance.

Deep grooveball bearing

Cylindricalroller bearing

Fig. 1 Single-layer, ceramic-coated, insulated bearings*Industrial Sales Headquarters Industrial Engineering Department-48-

Insulated bearing "MEGAOHM" series(Test Results)

2. Features of Single-Layer, CeramicCoated, Insulated Bearings

A slight indentation or mark was observed in thearea exposed to impact, but the insulation resistanceafter immersion in hot water remained 1000 MΩ orhigher.

1. Available as a drop-in replacement for standard
bearing sizes
2. Applicable bearing types include deep groove ballbearings and cylindrical roller bearings
3. Insulation performance
Insulation resistance: ≧100MΩ
(500V forced, at 20˚C)
Dielectric breakdown voltage: ≧2kV (AC60Hz)

3. 2 Mounting and Dismounting Test
In some applications, the bearings need to bedismounted and re-mounted repeatedly for periodic
inspection. A mount/dismount test was conducted tocheck whether damage (if any) to the ceramic-coating
would result in deteriorated performance.

3. Evaluation Tests for Single-Layer,
Ceramic-Coated, Insulated Bearings(Test Method)
The equipment shown in Fig. 4 was used todismount and re-mount the bearing five times. Then,the bearing was immersed in hot water, and thechange in insulation resistance was measured tocheck for the presence of abnormalities in the plasmacoated ceramic layer. In an actual application, theceramic surface should be coated with used grease.
However, this test was conducted under a severe
condition with no grease coating.
The fit between the outer diameter of the outer ringand inner diameter of the equipment is φ170 (36μmtight fit).

3. 1 Drop Test
A drop test was conducted to determine the effect ofan impact load on the insulated surfaces.

(Test Method)
A 6316 bearing was dropped from a height of 50
mm as shown in Fig. 2, and then immersed in hot
water (80˚C for one hour). If cracks are present, water
passes through , resulting in decreased insulation
resistance. The insulation resistance was measured
using the equipment shown in Fig. 3.

(Test Results)

50mm

After the test, no abnormalities were found on theceramic-coated surfaces. In addition, it was confirmedthat the insulation resistance after immersion in hot
water was again 1000 MΩor higher and the ceramiccoated layers were not damaged.

Press-fit (36 μm tight fit)

Metal plate

Fig. 2 Schematic of drop test

Insulation resistancetesterFig. 4 Mounting and dismounting test

Fig. 3 Measurement of insulation resistance-49-

NTN TECHNICAL REVIEW No.71（2004）

3. 3 Humidity Test

(Test Results)

A test was conducted to check the effects oftemperature and humidity on the insulation
performance. Typical environmental conditions wereused during this test.

After the test, no abnormalities were found on theceramic-coated surface . In addition, it was confirmedthat the insulation resistance after immersion in hot
water was 1000MΩ or higher and the ceramic-coatedlayers were not damaged.

(Test Method)
3. 5 Rotation Performance Test

The bearings were placed inside thetemperature/humidity-controlled chamber andexposed to the following conditions for 10 cycles.

Since the ceramic-coated surfaces have a lower
thermal conductivity than bearing steels, the heatradiation property was checked while rotating thebearings.

Temperature 40˚C, humidity 85%: 2 hours
1 cycleTemperature 25˚C, humidity 50%: 0.5 hours(Test Method)

Four bearings were tested.

Condition 1: Test bearing part no.: 6316
Test equipment: See Fig 5.
Radial load: 3920N
Rotating speed: 1000, 2000, 3000, 4000, 5000 min-1
Grease: Unimax R No.2
Grease Fill: 30% of free space(Test Results)
The insulation resistance was measured
immediately after the 10-cycle test. This test revealed
that the bearings, which had an insulation resistanceof 1000 MΩ or higher before the test, provided aminimum of 180 MΩ resistance. After the bearings
were left to stand, the insulation resistance wasmeasured again. At this time, all the bearings showed1000 MΩ or higher. This indicates that the ceramiccoated layers developed condensation when exposed
to high humidity. So, when handling bearings, caremust be taken to avoid humid environments.

Condition 2: Rotating speed 4000 min-1 (endurance
test)* All other conditions are the same as in Condition 1 above.

(Test Results)
Table 1 shows the temperatures measured at eachof the rotating speeds defined in Condition 1.
Table 2 shows the temperatures measured after
operating for 2,200 hours as defined in Condition 2.

3. 4 Heat Deterioration Test
A test was conducted to consider the effects of
rapid temperature change on the insulation
performance. Temperatures were based on typicaloperating conditions.

Table 1 Temperature rise test (˚C)
Rotating speed (min-1)
1000
Surface temperature ofouter ring (bearing steel) (1) 31.9

(Test Method)
Condition 1: The bearings were placed inside thethermal shock chamber and exposed to thefollowing conditions for 20 cycles.

2000

3000

4000

5000

33.8

37.2

39.6

43.5

Surface temperature ofceramic layers (2)

31.6

33.5

37.0

39.4

43.3

(1) - (2)Inner ring temperature

0.3
31.3

0.3
36.5

0.2
41.3

0.2
44.8

0.2
49.7

* The temperatures shown at each rotating speed were those when
the difference in surface temperature between outer ring (bearing
steel) and ceramic layers was the largest.

1hr

130˚C
0.5hr

1hr

0.5hr

Normaltemperature

Table 2 Bearing endurance test (˚C)

hour

Test time (hr)Surface temperature of outer ring (bearing steel) (1)Surface temperature of ceramic layers (2)(1) - (2)Inner ring temperature-30˚C

1 cycleCondition 2: The bearings were placed inside thetemperature-controlled chamber and left at
160C˚ for 2,200 hours.

-50-

2200
39.2
39.2
0
43.9

Insulated bearing "MEGAOHM" series¡Measuring point for surface temperature of ceramic layer¡Measuring point for outer ring temperature※The outer ring temperature was measured with the ceramiclayer removed from the area 45 degrees from the measuring
point for surface temperature of ceramic layers.
Support bearing¡Measuring point for
inner ring temperature

Test bearing (6316)

Fr
Fig. 5 Test rig

4. Conclusion
This paper introduced single-layer, ceramic-coated,insulated bearings that can be used for wind turbine
generators. Tests conducted under various operating
conditions showed high insulation levels of ceramiclayers. The rotation test also showed only a small
temperature difference of 1˚C or less between theouter ring material and ceramic surface, indicating that
thermal conduction was sufficient and rotation
performance was satisfactory. This newly developed
bearing will provide sufficient prevention of electric
corrosion in most applications that currently
experience such problems.

Photo of the author

Hideji ITO
Industrial Engineering Department
Industrial Sales Headquarters-51-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]New High-Capacity HWTJ Type Pressed Cage
and Needle Roller Assemblies

Katsufumi ABE*

Needle roller bearings offer compact size, large basic load ratings, and high rigidity compared to ball bearings.
Because of these properties, the use of cage and needle roller assemblies assists our customers with size and weight
reduction of their products.
In recent years, customers have called for bearings that have ever-increasing capacity. In response, NTN created a
new high-capacity pressed cage and needle roller assembly as the HWTJ Type. Compared to the standard design, thenumber of rollers was increased by 20 to 30% and basic load rating by 20 to 35%. This bearing series satisfies therequirements of large basic load rating, long life under severe operating conditions, and high rigidity.
This report introduces the structure and the performance of the new high-capacity HWTJ Type pressed cage andneedle roller assembly.

1. Introduction
Compared to ball bearings, needle bearings not
only require smaller space but also offer larger load
capacity and higher rigidity. Cage and needle roller
assemblies are composed of only a cage and a set ofrollers. They do not have inner or outer rings and use
the shaft and housing as the raceway surfaces.
Because of this, they enable the smallest and lightestweight design among the needle bearings, making it
possible to reduce the size of some machines in which
they are used.
Cage and needle roller assemblies are used as
bearings in the track planetary reduction gears ofconstruction machines, swing arm planetary reduction
gears of hydraulic excavators and in the arm joints ofindustrial robots. Recently, as a result of the drastic

reduction in machine size and advancement offunctions, a bearing with a larger load capacity (sizereduction) is required. Currently, in these applications,M-shaped cages (PK type) are often used. (See Fig. 1
and Photo 1.)
NTN has developed a new high-capacity pressed
cage and needle roller assembly (HWTJ type) for thismarket, featuring a larger number of rollers whencompared to the M-shaped cage. (See Fig. 2 andPhoto 2.)This bearing series also offers a drastically
improved basic load rating, as well as longer life andhigher rigidity under severe operating conditions
without sacrificing necessary bearing features (i.e.
smooth rotational movement, durability).

*Needle Roller Bearing Engineering Department Automotive Sales Headquarters-52-

New High-Capacity HWTJ Type Pressed Cage and Needle Roller Assemblies

A

A

B

Outer pillar

Outer pillar

B

Inner pillar

Inner pillar (Retaining ring)
B-B view

A-A view

Fig. 1 PK typeFig. 2 HWTJ type

Photo 1 PK type

Photo 2 HWTJ typeThis has not only increased the load capacity (basic
static load rating) by 20 to 35%, but also improved
bearing rigidity and life drastically. Table 1 shows acomparison of the bearing specifications between the
new roller assembly and the conventional type (PKtype), and Table 2 shows the comparison of theirgeneral performance.

2. Features of New High-Capacity(HWTJ) Type Pressed Cage andNeedle Roller Assemblies
The only way to achieve higher load capacity forcage and needle roller assemblies that consist of only
two components (cage and roller) is to increase thenumber of rollers. With the conventional M-shaped
cage (Fig. 1), pillars are located at the inner diameterside to prevent the rollers from dropping out. Thisrequires the cage to pass through the bearing PCD,
thereby limiting the number of rollers that can beinserted in a given bearing size.
To solve the conflicting issues of "higher load
capacity" and "roller drop-out prevention", the highcapacity pressed cage and needle roller assembly
consists of three components (a cage, rollers, and aresin retaining ring). The form of the cage was alsochanged and the method of retaining the rollers was
modified to include a resin retaining ring, thus
eliminating cage pillars at the bearing PCD. This
made it possible to increase the number of rollers by
20 to 30% when compared to the conventional
product (M-shaped cage).

Table 1 Comparison of bearing specifications(Size :ID46×OD66×Width 22.8 mm)
Developed product Conventional product(HWTJ type)(PK type)
Number of rollers

16

12

Roller length (mm)1818Basic dynamic load rating (kN)

84.0

67.5

Basic static load rating (kN)

98.5

73.5

Table 2 Comparison of general performance

-53-

Item

Compared to conventional type

1 Number of rollers
2 Static load rating (impact load)
3 Dynamic load rating
4 Calculated bearing life
5 Rigidity

Increased by 20 to 30%
Increased by 20 to 35%
Increased by 15 to 25%
1.5 to 2 times
Increased by 15 to 25%

NTN TECHNICAL REVIEW No.71（2004）

4. 1. 2 Test Conditions

3. Main Features of New Technological
Development

Radial load: 37.7kN : (45% of dynamic load rating Cr)
Rotating speed : 1,560 rpm
Lubrication : Grease
Calculated L10 life : 152 hours(Reference : The same size PK type bearing had alife of 59.8 hours)

Items to note about the design of this bearing aresummarized in the following two points.
1 Roller Retention
Resin retaining rings are provided on the inner sideof the cage to prevent rollers from dropping out
inwards.
2 High load capacity (increased number of rollers)
The positions of the roller retaining pillars weremoved closer to the inner and outer diameter of the
bearing to eliminate the pillars near the PCD thatwere limiting the number of rollers that could be
used.

4. 1. 3 Test Results
Fig. 4 shows the results of the bearing life test. The
major damage observed was flaking of the shaft. The
bearing life was 178 hours, which exceeded thecalculated L10 life of 152 hours.

99

Cumulative percent failed (%)

4. Performance Evaluation
In order to evaluate the performance of the
bearings, a life test and a strength test for the cageand retaining rings were conducted. Some of theseevaluation results are introduced below.

4. 1 Bearing Life Test
A 2 to 4-ton radial load test machine shown in Fig. 3
was used to conduct the bearing life test.

80
50
20
10
5

1
101

102

103

Fatigue Life (hour)

4. 1. 1 Bearing Specifications

Fig. 4 Rolling contact fatigue life of bearing

Bearing size :ID46×OD66×Width 22.8 mm
Roller size :OD10×length 18 mm
Number of rollers : 16

Loading lever

Test bearing

Fig. 3 2 to 4-ton radial load test machine

-54-

104

New High-Capacity HWTJ Type Pressed Cage and Needle Roller Assemblies

4. 2 Confirming Roller Interaction with Cage andRetaining Rings

4. 3. 1 Calculation Conditions
Bearing size :ID46×OD66×Width 22.8 mm
Roller size: OD10×length 18 mm
Number of rollers : Developed bearing: 16,
Conventional bearing: 12
Load: 1, 5, 10, 50, 100 kN

One of the applications for which this bearing canbe used is a planetary reduction gear used inconstruction machines. In these machines, thebearing rotates with oscillating motion. Unlike the Mshaped cage that has pillars near PCD to guide therollers, this bearing has cage pillars near the bearing's
outer diameter and retaining ring pillars near the
bearing's inner diameter. To ensure that the rollers do
not run onto the pillars during changes in rotation
direction, a test was performed under the oscillation
conditions listed in 4.2.1. For this test, a bearing
having the same specifications as the life test wasused in a cross-joint endurance test machine. The
structure of the bearing mating components in thismachine is the same as the 2 to 4-ton radial load test
machine shown in Fig. 3.

4. 3. 2

Calculation Results

Fig. 5 shows the calculated rigidity comparison. The
rigidity of the developed bearing (HWTJ type) is
approx. 20% higher than that of the conventional
bearing's rigidity.

Bearing rigidity (×106 N/mm)

1.8

4. 2. 1 Test Conditions
Radial load: 37.7kN : (45% of dynamic load rating Cr)
Oscillation angle: ±720˚
Oscillation cycle speed: 48 cpm
Lubrication: Grease
Test duration: 1 million cycles

Developed bearing

1.6

Conventional bearing

1.4

1.2

1.0

0.8
0.6
0.1

4. 2. 2 Test results

1

10

100

1000

Load (×103N)

Table 3 shows the test results. This data shows thatthe bearing operated more than one million cycles,
which is the NTN test standard, without any issues
with the rollers running onto the cage pillars or
retaining ring pillars.

Fig. 5 Comparison of rigidity

5. Conclusion
Table 3 Test results
Number ofcycles1 million
No.1 suspend

Cage/retaining ring
No problem found withthese components

No.2 1.72 million No problem found withcyclesBy designing and building a bearing using threecomponents (cage, rollers and resin retaining ring)
and adopting a cage drawn from a thick plate, wehave succeeded in the development of "New HighCapacity Pressed Cage and Needle Roller
Assemblies" featuring a high load carrying capacity
and compact size.
We will continue to work on the development of
products such as this that offer larger load capacities
and a more compact design.

Appearance after test

these components

Roller/shaft/outer ring
No Problem.
Flaking on outer ring
No problem found withrollers and shaft

4. 3 Bearing Rigidity
This new bearing can have 20 to 30% more rollers
than the conventional bearing, without any
modification of the major envelope dimensions. This
has enabled the use of the HWTJ type bearing inapplications where the conventional bearing was notadequate. In addition, since the HWTJ bearing is morerigid than the conventional bearing, under the sameload products using the HWTJ bearing can havehigher precision.
To compare bearing rigidity with the conventional
bearing, a NTN technical computing program was used.

Photo of the author

Katsufumi ABE
Needle Roller Bearing Engineering Department
Automotive Sales Headquarters-55-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]HK-F type drawn cup needle roller bearings

Hideki AKAMATSU*

The drawn cup needle roller bearing has an outer ring which is precisely drawn from a thin steel plate. Of all bearings
with outer rings, drawn cup needle roller bearings have the smallest section height, which enables space and costsaving.
NTN has applied this technology to many applications.
NTN's next generation "HK-F Series" are standard drawn cup needle roller bearings to which NTN has applied newtechnology.

1. Introduction
Since needle roller bearings have a small section
height and provide larger load capacity and rigidity
than the other type bearings in spite of small space,
they are used in many applications including theautomotive field. In particular, drawn cup needle roller
bearings (hereafter called drawn cup NRBs) have anouter ring that is drawn from a steel plate. In addition
to having the lowest section height of any needle roller
bearings with outer rings, they have excellent cost
performance. Furthermore, because they use a pressfitted housing, they do not need a snap ring or otherdevice to fix them in the axial direction. This expands
the range of applications to include those of
conventional machined-ring needle roller bearings.

Special specifications have been employed to thestandard type bearings in response to the increasing
demand for longer life, reduced fuel cost, and greater
power particularly in the automotive field.
Some of these special specifications have been
employed as standard specifications in order to
respond to worldwide needs. This has enabled NTN tooffer its standard drawn cup bearings "HK-F Series",featuring longer life, improved load resistance, andeasier assembly.
Now, I will introduce its features, design concept,
and evaluate test results.

*Needle Roller Bearing Engineering Department Automotive Sales Headquarters-56-

HK-F type drawn cup needle roller bearings

2. Features
The features of the HK-F type drawn cup NRB aregiven below (Fig. 1)).

■JIS SCM steel used for outerring■Unified cage type

Result: improved lubrication( V-form welded cage for Fw ≥φ5mm;
plastic cage for Fw < φ5mm)

Result: increased casehardness■Optimal outer ring shapeincluding chamfer and outer
diameter surface

Result: easier assembly■Soft nitrided, or case hardened
and tempered heat treatment

Result: improved cage wear

■Roller crowning added toroller outer diameter

Result: reduced edge
stress, longer life
■Increased inner diameter ofouter ring bent edge

■Special heat treatment forrollerResult: better lubrication

Result: longer life

■Wider use of "Pre-bent" method

Result: free directional fitting,longer life

Fig. 1 Specification of HK-F type drawn cup NRB

2. 1 Long Life
Contact surface pressure

HK-F type drawn cup NRB has a longer life as aresult of the following improvements.

1 Adoption of crowning as a standard feature
Stress concentration (edge load) at the roller ends
caused by assembly misalignment and heavy load
is reduced, enabling use of the standard typeunder a wider range of conditions. Fig. 2 shows
the calculation results of contact surface pressure
between the current standard type and HK-F type.
2 Adoption of a special heat treatment on rollers
as the standard featureSpecial carbonitriding improves resistance tosoftening that occurs as a result of tempering andincreases the amount of residual austenite, whicheases stress concentration.
This has enabled HK-F type to have a threetimes longer life than the current standard typewhen used under clean lubrication conditions(Figs. 3, 4)).
3 Adoption of "Pre-bent" method (Only for some
HK-F type drawn cup NRBs)
Bearings manufactured by "Pre-bent" method(patent obtained, explained later) provide evenlonger life as a result of optimization of heat
treatment conditions.

Current standard product

Effective roller length

Contact surface pressure

No crowning

No misalignment

HK-F type

Effective roller length

Contact surface pressureWith crowning

No misalignment

HK-F type

Effective roller lengthWith crowningWith misalignment

Fig. 2 Calculation result of contact surface pressure-57-

NTN TECHNICAL REVIEW No.71（2004）

Test bearing
Oil supply
Oil drain
Inner ring IR17×22×26

Cumulative percent failed (%)

Fig. 3 Bearing life test rig

Current standard product
HK-F type

95
90
80
70
60
50
40
30

Test bearing: (Current) HK2216C (HK-F type) HK2216F
Load capacity: (Current) 13.6kN (HK-F type) 13.2kN
Test load: 6.82kN《approx. 0.5˚C》
Rotating speed: 10000 min-1
Lubricating oil: Turbine oil 46
Lubrication system: Circulating oil

20
10
5

2

3

5

7

1
1
10

2

3

5

7

1
102

2

3

Life (h)Fig. 4 Bearing life test results

4 Adoption of V-form welded cage (Onlyfor some HK-F type drawn cup NRBs)
Adoption of the V-form welded cage
design allows use of roller guide closer tothe roller center than the conventional Uform cage, as a result movement of roller
during rotation is smooth. Furthermore,
the cage rib can be shaped so as toprevent flow-in/-out of lubricating oil,resulting in better lubrication and better
effect on the sliding surface.
Fig. 5 shows a cross-sectional view of
the current and HK-F type drawn cup
NRB cages.
For heat treatment of cage, soft nitriding(carburization for "Pre-bent" type) has
been adopted as the standard feature to
improve wear resistance and strength.

An example of current standard typeU-form pressed cage

HK-F type
V-form welded cage

Fig. 5 Cross section of current and HK-F type drawn cup NRB cages-58-

HK-F type drawn cup needle roller bearings

2. 2 Improvement of Allowable Static Load

2. 3 Reduction of Initial Mounting Force andStabilization of Mounting Force

Chrome molybdenum steel (JIS SCM steel) has
been adopted as the material for the drawn cup.

1 Improved assemblyThis has increased inner hardness of the drawn
cup, widening the usage range for heavy-load
applications.
Fig. 6 shows the method of static load test
conducted to compare HK-F type with the current
standard type, and the results are shown in Fig. 7.
Since HK-F type drawn cup NRBs do not develop
permanent deformation even at loads of S0 ≧2 (upto 50% of static rated load) while the current
standard bearings can be used at S0 ≧3.

Drawn cup NRBs are press-fitted into the
housing.
With HK-F type drawn cup NRB, outer diameterchamfer shape is optimized to reduce the initial
mounting force and stabilize the mounting force.
Fig. 8 shows the results of the fit test.

10

3 rollers(φ3×7.8mm)

Backup
JIS SUJ2 steel
through hardened

Test ring

Total permanent deformation (μm)

5
4
3
2

1.8
0.5

0

2

4

6
8
10
Press-fit amount (mm)1214Bearing: (Current) HK2216C(n=2)(HK-F type) HK2216F
Material of housing : steel
Fit: Recommended value given instandard catalog

Current (JIS SPC steel)
HK-F type(JIS SCM steel)

15

6

0

20

16

7

1

Fig. 6 Static load test condition

18

8
Initial mounting force

Press-fit mounting force (kN)LoadThree rollers
equally spacedin cage

Current standard product
HK-F type9Note) S0 (safety factor) = Static rated load / Radial load

Fig. 8 Fit test result14

2 Free Direction Fitting ( "Pre-bent" method)

12

In the manufacturing process of conventionaldrawn cup bearings, the rib of the heat-treatedouter ring (at the bent edge side) is annealed first,
and then the rollers and cage are assembled into
the outer ring. Finally, the rib is bent to prevent
separation. (Fig. 9) As a result, the hardness of therib (bent edge side) is low, therefore, press-fitting
the bearing into the housing by pushing on thebent edge rib is not desirable in terms of strength.
With drawn cup NRBs of "Pre-bent" method, therollers and cage are assembled into the outer ringfirst, and then the bent edge rib is bent prior to heat
treatment. This gives the rib (at the bent edge side)
the same hardness as that of the other areas. As aresult, the hardness of the rib (at the bent edgeside) increases compared with conventional cup
bearings, enabling free directional fitting into the
housing.

10
8
6
4
1.4
4
0

0.43

1

0

2
Conditions

Material of rings: (Current) JIS SPC carburized steel(HK-F type) JIS SCM steel carburized steel
Ring Dimensions: Thickness 1mm, flat
Roller: φ3×7.8mm, 3 rollers, crowning
Test load: 1 9.74kN 27.31kN
Pmax: 13.9GPa 23.4GPa
S0 converted: 11.5
22.0

Fig. 7 Comparison of static load test results-59-

NTN TECHNICAL REVIEW No.71（2004）

Heat treatment of outer ring↓Annealing of bent edge side↓
Assembly of rollers and cage↓
Bent edge rib bending

Assembly of rollers and cage
into non-heat-treated outer ring↓
Bent edge rib bending↓
Heat treatment of assembly

Annealing(1) Standard method(2) "Pre-bent" method

Fig. 9 Process of bent edge rib bending

Table 1 NTN HK-F type drawn cup NRBs products
Boundary dimensions (mm)Inscribed
Outer
Width
circlediameter diameter
D
B
Fw
6
6.5
3
8
8
4
9
9
5
9
10
6
9
11
7
10
12
8
10
13
9
12
13
10
14
12
14
10
15
14
10
16
12
12
18
12
19
13
12
20
14
16
20
12
21
16
15
21
22
21
12
22
16
22
16
22
22
12
23
17
12
24
18
16
24
12
26
16
26
20
20
26
30
26
12
28
16
28
22
20
28

Bearing numbers
Current catalog
part number
HK0306T2
HK0408T2
HK0509T2
HK0609T2
HK0709HK0810C
HK0910HK0912
HK1010HK1012
HK1015
HK1210HK1212
HK1312
HK1412
HK1416
HK1512
HK1516
HK1522ZWD
HK1612
HK1616
HK1622ZWD
HK1712
HK1812
HK1816HK2012C
HK2016
HK2020C
HK2030ZWD
HK2212HK2216C
HK2220C-60-

HK-F type part number
HK0306FT2
HK0408FT2
HK0509FM
HK0609FM
HK0709FM
HK0810FM
HK0910FM
HK0912F
HK1010FM
HK1012F
HK1015F
HK1210FM
HK1212FM
HK1312FM
HK1412FM
HK1416F
HK1512FM
HK1516F
HK1522ZWFD
HK1612FM
HK1616F
HK1622ZWFD
HK1712FM
HK1812FM
HK1816F
HK2012FM
HK2016F
HK2020F
HK2030ZWFD
HK2212FM
HK2216F
HK2220F

Polyamide Resin cage
Polyamide Resin cage
"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method"Pre-bent" method

HK-F type drawn cup needle roller bearings

3. Conclusion
The HK-F type drawn cup NRBs have a longer life
and improved functionality compared with the current
standard type, which enables its use in applications
where special specifications are required.
Currently, a total of 32 HK series models, ranging
from φ3mm to φ22mm in inscribed circle diameterare available (see table 1). The application range ofthis specification will be expanded in the future.

Photo of the author

Hideki AKAMATSU
Needle Roller Bearing Engineering
Department
Automotive Sales Headquarters-61-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]Hydrodynamic BEARPHITE Unit for HDD

Kiyotaka KUSUNOKI*

Hard disk drives are used mainly as storage systems in personal computers, and it is expected that consumer usage
will continue to grow in the near future. In order to increase the storage capacity on hard disk drives, the Non-repetitive
Run Out (NRRO) needs to be decreased. In addition, requirements for shock endurance and low noise have led to therapidly increasing use of hydrodynamic bearings for hard disk drive spindles.
Hydrodynamic bearings are suitable because of their compact size and rotational accuracy.
NTN has long been working on the hydrodynamic bearings to solve these HDD problems, and mass production of
the BEARPHITE unit began last year. The details of this bearing are explained below.

1. Introduction
The surface recording density of hard disk drives(HDD) is doubling every year, and an HDD with asurface recording density of 80GB per disk has beenreleased to the market. Among the many storage disk
alternatives, HDD has established itself as the leader
because of its low cost per storage capacity unit andhigh data transfer speed. Conventionally, HDD wereused primarily in personal computers and servers.
However, use in consumer products (recorders, car
GPS navigators, video game consoles, etc.) has
already begun and this market is expected to continue
to grow.

For HDD spindles, rolling bearings have typically
been used. However, in order to improve the Nonrepetitive Runout (NRRO) and meet requirements forhigh shock endurance and low noise, the use ofhydrodynamic bearings for HDD spindles has been
rapidly increasing over the last two years.
NTN has worked on the development ofhydrodynamic bearings for many years and solved
several technical problems. NTN has begun mass
production of hydrodynamic BEARPHITE units
equipped with porous oil retaining bearings. An
overview of the hydrodynamic BEARPHITE unit isgiven below.

*Fluid Dynamic Bearing Unit Division-62-

Hydrodynamic BEARPHITE Unit for HDD

strong demand for HDD spindles with improved
NRRO. Manufacturers are switching from ballbearings to hydrodynamic bearings since quietness
and shock resistance are required for the application.
Rolling bearings are at a disadvantage because theirrolling elements contact the raceway, thereby placing
a limit on the size of the track pitch even if the
raceway is machined very precisely.
Table 1 shows a summary of requirements for HDD
and the characteristics of various types of bearings.
From a functionality comparison, hydrodynamic
bearings (including hydrodynamic BEARPHITE unit)
are better for NRRO, high-speed rotation, noise leveland shock resistance, but worse for rigidity andreliability. Both bearings showed about the same
torque. Porous bearing sleeves are anotheradvantage of hydrodynamic BEARPHITE units,
because lubricating oil is retained inside the sleeves
and is supplied to the bearing continuously. This
prevents bearing seizure, thereby allowing theirreliability to exceed that of standard solid metalhydrodynamic bearings.

2. Classification and Features of
Bearings
2. 1 HDD and Bearing Characteristics
Rolling bearings that are designed solely for
improvement of NRRO have been used in HDD
spindles because of their price, easy handling andreliability.
Hydrodynamic bearings retain oil between the shaft
and bearing in special grooves. These grooves
produce a high oil film pressure, which enables thebearing to be supported without being in contact withthe shaft. Because of the oil film, imperfections in theshaft surface condition (roundness, surface
roughness) and bearing are diminished, providing highrotational accuracy (NRRO). Therefore,hydrodynamic bearings have been determined to be
more suitable for HDD than roller bearings due to thecompactness and high rotational accuracy required.
In order to use hydrodynamic bearings in HDD,
electric-appliance and bearing manufacturers havespent much time on the development of thesebearings. Since hydrodynamic bearings require bettercomponent precision and precise machining of thehydrodynamic grooves, they are more expensive thanball bearings and suffer from reliability problems due
to an abrupt seizure known as sudden death.
Because of these reasons, they are not very common.
Over the past decade, the track pitch of HDD units
has been reduced to increase the storage capacityand now can be measured in tens of nanometers.
The head can respond to low-frequency repetitive
runout (RRO) but cannot follow irregular-frequency
non-repetitive runout (NRRO). Therefore, there is a

2. 2 Lubricating Oil

Lubricating oil plays an important role inhydrodynamic bearings since the function equivalent
to rolling bearings is required. The lubricating oil
requires the following characteristics for long life andhigh performance. Recently, ester-based lubricating
oil has been used in hydrodynamic bearings.
1 Excellent lubricating performance: Improved
contact condition between shaft and bearing at
start and stop.
2 Sealing performance: Prevent leakage from thebearing due to higher surface tension and lower
specific gravity.
Table 1 Requirements and characteristics
3 High temperature-viscosity◎ : Very good ○ : Good △ : Fair
Requirements
Requirements
Roller Hydrodynamic Hydrodynamic
characteristic: Reduce thefor HDDfor bearing
bearing bearing ＊ BEARPHITE unit ＊ Typical applications
changes in torque and
NRRO
△
○
○
Personal computer, server
Large storage
rigidity caused bycapacity
High-speed rotation△○○
Servertemperature change.
Low noise level
Low noise level○◎◎
Consumer
4 Low evaporation: Prevent○○○
MobileLow current consumption Low torqueImpact resistance Impact resistance○◎
Mobilethe reduction of lubricating
Reliability
Seizure resistance◎△○
Generaloil amount even if used for
Rigidity
Rigidity◎△△Serverlong periods of time.
＊The hydrodynamic bearing is made of solid metal and the hydrodynamic BEARPHITE unit is
5 Low torque: Low viscosity
made of porous material.
with acceptable bearing
Table 2 Bearing specifications
stiffness
Current spec.→Required spec.
Remarks
0.05μm → 0.03μm
Storage capacity 40→80→120GB
High-speed rotation 10,000min-1 → 15,000min-1
20,000min-1 (for server)
Low noise level
30dB → 20dB
Low torque5mN･m → 3.5mN･m（-10˚C）Notebook computers, mobile devices with 2.5-in. or smaller hard disk
Notebook computers, mobile devices with 2.5-in. or smaller hard disk
Impact resistance 300G → 1,000G
Item

Radial NRRO

Seizure resistanceDesigned life: 7 years, Start/stop: 100,000 times-63-

2. 3 Required Specifications
Table 2 shows an exampleof the required specifications of
HDD spindle bearings.

NTN TECHNICAL REVIEW No.71（2004）

3. Structure
3. 1 Structure of Bearing Unitsurface tension in order to prevent leakage. When the
bearing is rotating, the lubricating oil is drawn towards
the bearing by centrifugal force to prevent splashing.
As the shaft rotates, the hydrodynamic slots pump
the lubricating oil and a high-pressure oil film develops(see Fig. 2), causing the shaft to float. Since the shaft
is rotating, when radial load is applied to the shaft, theminimum clearance between the shaft and thecylindrical bearing is off-set from the centerline of theforce.

Fig. 1 shows the structure of a hydrodynamic
BEARPHITE unit designed for HDD spindles. The
unit consists of four components: a shaft, ahydrodynamic BEARPHITE bearing sleeve (hereafter
called bearing sleeve), a housing and a thrust
bushing.
The shaft has two cylindrical sections that aredesigned to support radial loads from the bearing
cylindrical components. Hydrodynamic herringboneshaped grooves are located at these same two points
on the bearing sleeve bore. The radial clearance
between the shaft and the cylindrical bearing
components is on the order of microns. To prevent
negative pressure from being generated inside thebearing, the width of the herringbone grooves isclosely controlled.
The bearing's components are made so that theflange on the end of the shaft is sandwiched between
the cylindrical sleeve and the thrust bushing. The
thrust bushing has spiral-shaped thrust hydrodynamic
grooves. The flange, cylindrical sleeve and thrust
bushing are separated by clearances of just microns.
A seal is located at the point where the unit's shaft
extends from the housing to protect contaminationfrom entering the unit. The shaft outer diametersurface and housing inner diameter surface aretapered on the inboard side of the seal. When the
bearing is not rotating, the lubricating oil is retained by

3. 2 Structure of HDD
Fig. 3 shows the structure of a HDD. A rotor
magnet is attached to a hub and the hub assembly ispress-fit onto the shaft of a hydrodynamic
BEARPHITE unit. The HD spindle is formed when ahousing outer diameter surface is connected to abracket where a stator coil and control PCB are
attached. The actuator (pivot bearing with voice coilmotor) supporting the head is attached to the base,and it writes/reads signals while detecting recording
positions on the hard disk.

Flanged shaft
Seal

Bearing

Shaft: Off-centerand floating

Hydrodynamic
BEARPHITE
bearing sleeve

Housing

Thrustbushing

Radial bearing

Thrust bearing

Bearing
clearance

Pressure distribution
by hydrodynamic action

Formation of oil filmby hydrodynamic
actionAverageeffectiveness:
Smoothes out theshaft and bearing
imperfections
Constant torquebecause of noncontact between
shaft and bearings

High rotational
accuracy

Fig. 1 Structure of hydrodynamic BEARPHITE unit
(longitudinal cross section)

Fig. 2 Structure of hydrodynamic BEARPHITE unit
(radial cross section)-64-

Hydrodynamic BEARPHITE Unit for HDD

Cover
Head

Actuator
Hard disk
Write/read preamplifier
HDD spindle motor
Base

Insulator
Printed wiring board (PWB)

Fig. 3 Structure of HDD

4. Theory and Measured Values]
4. 1 Theoretical Calculation
With the porosity of the BEARPHITE sleeve and the
lubricating oil between the radial and thrust bearings
taken into account, a program was created utilizing
the narrow groove theory to analyze the performanceof the hydrodynamic BEARPHITE unit. This program
calculates the bearing pressure distribution anddetermines bearing stiffness, load capacity, torqueand attenuation factors. Use of this program helps
optimize HDD spindle rigidity, load capacity andcritical speed. Examples of bearing pressure
distribution calculations are shown in Fig. 4.

P, atmz, mm

θ, deg.

Fig. 4-2 Bearing pressure distribution(radial clearance)

P, atm
P, atm

r, mm

z, mm

r, mm

r, mm

Fig. 4-3 Bearing pressure distribution(axial clearance)

Fig. 4-1 Bearing pressure distribution(sleeve)-65-

NTN TECHNICAL REVIEW No.71（2004）

4. 2 Measurement Results(1) Radial stiffness
Fig. 5 shows the test equipment used tomeasure the radial stiffness of hydrodynamic
BEARPHITE units. A disk with weight (inertia)
equivalent to the actual machine was placed on
the shaft and driven by an air turbine. A magnet,
calibrated for distance and draw-in load, wasplaced close to the magnetic ring on the disk outerface. A radial load was exerted without any
contact between the two objects. The change ininertia was measured by two non-contact
electrostatic capacitance displacement gauges
positioned at a 90-degree phase (load/changeamount). The bearing stiffness value of the specimen
was measured to be 9.47×106 N/m.

Fig. 5 Radial stiffness measurement equipment(3) NRRO (Non-repetitive Run Out)
Fig. 8 shows the test equipment used to measure
the NRRO of hydrodynamic BEARPHITE units. The
hydrodynamic BEARPHITE unit was placed on a
motor and an aluminum disk with weight (inertia)
equivalent to the hard disk was attached to the unit.
Next the motor was turned on to drive the unit. Runout in both radial and axial directions was measured
using a non-contact electrostatic capacitancedisplacement gauge and was analyzed by FFT toobtain NRRO. The NRRO measurement results areshown in Fig. 9. It shows that radial NRRO was 0.010
to 0.015μm at room temperature and 0.013 to 0.024
μm at 60˚C. Both of these measurements satisfy(2) Torque
A disk with weight (inertia) equivalent to the actualmachine was placed on the shaft of the hydrodynamic
BEARPHITE unit. The housing was secured to the
top of the static air bearing (see Figure 6). To drive
the disk, a turbine blew air onto the disk outer
diameter. A strain gauge measured the tangential
force of the string attached to the pulley. The
tangential force was converted into a torque by
multiplying it by the pulley radius. Fig. 7 shows thetorque result of 1.2 to 1.4 mN･m.
Speed sensor

Non-contact type displacementgauge for axial load

Loading disk

Motor
Dummy disk

Z-axis stage

Air nozzle

Non-contact
displacementgauge forradial load

Specimen

Z-axis stage

Specimen chuck

X-axis stage

X-axis stage

Hydrostatic
air bearing
Motor fixture

Strain gauge
converter

Fig. 8 NRRO measurement equipment

Fig. 6 Torque measurement test equipment

At room
temperature

0.05
0.04

NRRO，μm

Torque，mN・m

2.0

At 60˚C

1.5
1.0
0.5

0.03

Upper limit of specification

0.02
0.01

0.0
1

2

3

4

0.00

5

Specimen No.

1

2

3

4

Specimen No.

Fig. 7 Torque measurement test results

Fig. 9 NRRO measurement results

-66-

5

Hydrodynamic BEARPHITE Unit for HDD

NTN's requirements.
Theoretically, the hydrodynamic bearing run-out has
a repeatable component and a non-repeatable(NRRO) with each rotation. The causes of the bearing
NRRO could be air flow near the measuring points,
unsteady lubricating oil flow, motor vibration,
centrifugal force caused by fluctuation of rotating
speed, change of hydrodynamic effect or component
precision not dampened by oil film.

(5) Durability
A HDD can quickly become obsolete in terms of its
functionality due to frequent model changes.
However, it is essential to provide the designed
durability during the warranty period (7 to 10 years).
In addition, it is very important that they do not stop
during use, since critical data is saved in personal
computers and servers via HDD's.
Figs 11-1 and 11-2 show endurance test results
taken while a HDD bearing was started and stopped
frequently by a motor running at a high temperature(60˚C). It shows that there was no change in the
NRRO or the motor's current during the test and nodecrease in functionality of the HDD.
For home use, this sound level is lower than the
background noise.

(4) Low Noise Level
The noise measuring method for HDD is stipulated
by both JIS and ISO standards. With rolling bearings,
noise is generated from rolling contact between therolling elements and raceways as well as sliding
contact between the cage and the other components.
To the contrary, no noise is generated byhydrodynamic bearings, since their shaft and bearings
do not contact. Most of noise is produced by theelectro-magnetic sound from the motor and the harddisk wind noise. Fig. 10 shows a comparison of test
results from a roller bearing and a hydrodynamic
bearing. The noise level of the hydrodynamic
BEARPHITE unit was 22 dBA, which was 8 dBA lower
than the roller bearing, showing that the hydrodynamic
BEARPHITE unit operates more quietly.

0.25

NRRO（μm）

0.20
0.15
0.10

0.05
0.00
0
40

50

100

Running cycle

150

200

×104 cycles

Fig. 11-1 Endurance test (radial NRRO)

20

250
200

10

Current (mA)

Noise level dBA

30

0Hydrodynamic
BEARPHITE unit

Rolling bearing

150
100

50

Bearing0

Fig. 10 Noise test results

0

50

100

Running cycle

150

×104 cycles

Fig. 11-2 Endurance test (motor current)

-67-

200

NTN TECHNICAL REVIEW No.71（2004）(6) Impact resistanceresults. The figures show there was no change ineither torque or NRRO. Although it is known that thesurface pressure generated between the shaft andbearing is lower than bearings using steel balls, thistest confirms that the unit is resistant to impact.

When using HDD's in mobile devices such as
notebook computers, it is important that there is no
drop in the HDD's functionality. An impact test was
conducted on the bearing unit and its functionality wasevaluated. Tables 3-1 and 3-2 show the test
conditions and Figs. 12-1 and 12-2 show the test

Table 3-1 Impact test conditions

Table 3-2 Impact test conditions

Ambient temperature Room temperature (approx. 23 to 24˚C)
Unit direction
Upright, horizontal, upside down
Disk weight load
5ｇ1200 G
Impact load(0.2 msec)
3 cycles continuously for each direction
Impact cycle
Specimen
3 specimens

Measuring apparatus
NRRO, torque tester
Rotating speed
4000, 7000 min-1
Ambient temperature Room temperature (approx. 23 to 24˚C)
Unit direction
Upright
Loading disk weight
5ｇ
Measurement interval Approx. 4 min. (2 min. for each rotating speed)
Measurement properties
Radial NRRO, torque

0.25

0.25

4000min-1

0.15

Before test

After test

After test

0.20

NRRO，μm

Torque，mN・m

0.20

Before test

7000min-1

0.10

0.05

4000min-1

7000min-1

0.15

0.10

0.05

0.00

0.00
1

2

3

1

2

3

1

Specimen No.

2

3

1

2

3

Specimen No.

Fig. 12-1 Torque comparison before and after impact test

Fig. 12-2 NRRO comparison before and after impact test

-68-

Hydrodynamic BEARPHITE Unit for HDD

5. Future Developments
needs. Since hydrodynamic BEARPHITE units
employ press-formed hydrodynamic grooves andneed to be price-competitive, design will work closelywith manufacturing from the early stages ofdevelopment to produce low-cost bearing units. This
paper has only introduced the hydrodynamic
BEARPHITE unit that was developed for HDD.
However, it can also be used in optical disk drives
such as DVD, polygon scanners and fan motors.
The hydrodynamic BEARPHITE unit is a product
NTN has been working on for many years andcontains various bearing technologies. NTN is using abusiness model where technological development anddesign are performed in Japan and the products are
manufactured in factories overseas.
NTN will continue to promote the advantages ofhydrodynamic BEARPHITE units to our customers inorder to increase their sales. NTN expects these units
to become one of the mainstream bearings productsin the future. NTN will actively follow our customer's
ever-changing requirements to make thehydrodynamic BEARPHITE unit acceptable worldwide
for HDD devices and strive for a high-quality
reputation in the marketplace. NTN expects that thehydrodynamic BEARPHITE unit will become
widespread in various fields.

It is expected that sales of HDD's will continue togrow due to the popularity of personal computers.
However, it is also expected to spread to theconsumer product field, including HDD recorders,mobile storage devices, digital video cameras andmobile phones. One more area of potential in thenear future is home servers. Currently, 3.5-inch
HDD's are 80% of total sales, but 2.5-inch and 1-inch
HDD are expected to expand in the future with new
applications.
One of the technological innovations expected toappear soon is the perpendicular magnetic recording
method. If it goes into mass-production, the storage
capacity will increase further. As a major trend, therequirements for the bearing's rotational accuracy(such as RRO) will become more severe. Whenclassifying requirements by application, home
products will require low noise level, mobile products
will require low torque to reduce power consumption
and higher impact resistance, and navigation systems
will require a wider operating temperature range since
they are used in cars. On the other hand, there will
still be a need for low-cost products even though theirstorage capacity is limited.
NTN will continue to improve functionality andpromote product development to respond to these

Photo of the author

Kiyotaka KUSUNOKI
Fluid Dynamic Bearing Unit
Division-69-

NTN TECHNICAL REVIEW No.71（2004）[ New Product ]High Angle Active Link

Keisuke SONE*
Hiroshi ISOBE*
Koji YAMADA*NTN has developed a High Angle Active Link that can control movement with two degrees of freedom. Thisequipment consists of a constant velocity joint(CVJ) with a power unit. This system is expected to be used for robotic
joints and as an optical platform.
This paper introduces the advantages that the constant velocity joint has over a parallel link mechanism and someconsiderations on its link positions obtained by analytical study of ADAMS(a dynamic modeling software package).

1. Introduction
NTN is looking to improve joints with a linkmechanism, for example a constant velocity joint withhigh bend angle. Such a joint is composed of links
whose connection can rotate with one degree of
freedom. We have located actuators in the linkconnector as an improvement to these joints. As aresult, we developed a device (Photo 1) that can becontrolled with two rotational degrees of freedom
between input and output segments. We expect that
this device can be used in robotic joints, opticalplatforms and so on. It can also be used alone as amachine element in the connection of parts between
the segments.
This paper introduces a device, consisting of a joint(OD49×height 48mm) and two actuators in the input
segment.

Photo 1 High Angle Active Link

*New Product Planning Department Research & Development Center-70-

High Angle Active Link

link of each linkage system lie only on the planebisecting the angle between the center lines of input
and output-side link hub. Therefore, this mechanism
has the constant-velocity performance.

2. Structure and Features
2. 1 Joint Mechanism
As shown in Fig.1, the joint forms a parallel linkage
mechanism, consisting of three linkage systems (input
arm, intermediate link, output arm) between the input
segment (input-side link hub) and output segment(output-side link hub). Each linkage system includes
four rotation pairs and three links. A bearing is used ineach rotation pair to reduce rotational resistance andeliminate clearance in the linking areas. Also shown in
Fig. 1, each linkage system forms the spherical linkmechanism (i.e. segments move on spherical surface)and has the input and output segments of the sameform. The shafts extending from the input- and outputside arm links are connected at certain crossing angle(angle γ, Fig, 2) via the middle link. These linkage
systems are located symmetrically to Point A, thecross point of two connecting rods of the middle link
toward input- and output-side.
The three linkage systems mentioned above arespaced evenly about a circle, and are constrained tomove on a common circle. This means that the crosspoint A of the two connecting rods inside the middle

2. 2 Unit Configuration
As shown in Fig.3, two actuators and reduction
gears are located inside the input-side arm connectedto the input-side link hub, to control the rotation angleof the arm. Motor 1 is located in the 1st linkage
system, motor 2 in the 2nd linkage system, and the
3rd linkage system is used as a passive link (Fig. 1).
Because of the parallel structure, the output side ofeach linkage system moves in such way that it
maintains symmetrical position to the points on thebisecting plane. This allows positioning of the output
segment for two angles (bend angle θ. swing angle
φ, Fig. 2).
Fig. 4 shows the relationship between the input-side
arm swing angles (see Fig. 5) of the 1st or 2nd
linkage system and two angles of output segment.
Motor 1 and motor 2 control these input-side arms. It
shows that the input-side arms turn within ±45
degrees and the phase difference in the input-side
arm swing angle between 1st and 2nd linkage
systems is 120 degrees.

3rd linkage system

Output segment(Output-side link hub)

Joint assy
Weight:75gr

Output-side
arm link

Middle link

48mm

input-side
arm link

2nd linkage
systemInput segment(Input-side link hub)M2 (motor 2)

Gear linking
input-side arm
and reduction
gear

Joint

φ49mm

Reductiongear

1st linkage
system
M1 (motor 1)

Actuator
Servo motor

Fig. 1 Joint assy
Fig. 3 Unit
Bend angleθ

A
Crossing
angleγArm rotation angle (deg)Swing angleφ50

Center of output-side
spherical link

Output-side sphericalsurface link
Circular trajectory ofpoint A

Bisecting plane

M1-θ：45˚
M1-θ：90˚
M2-θ：45˚
M2-θ：90˚

40
30
20
10
0
-10

0

30

60

90

120

150

180

210

240

270

300

330

360

0˚ of arm rotation
angle indicates the
position at thebend angle 0˚

-20
-30
-40

Input-side sphericalsurface link

-50

Center of input-side
spherical link

Swing angle φ(deg)Fig. 4 The relationship between the swing angle (φ) andthe revolution angle of the arm in the position of the bend
angle θ=45˚ and θ=90˚

Fig. 2 Operating principle of the joint-71-

NTN TECHNICAL REVIEW No.71（2004）

2. 3 Features

φ is near 30 (φ＝60˚) and 270˚ as shown in Fig. 7. In
motor 2, the load has the peak when φ is near 30(φ＝0˚) and 150 as shown in Fig. 8. The positions ofpatterns A, B and C in Fig. 9 correspond to theseswing angles (φ=30˚, 150˚, 270˚). Note that thetorque at part 2 and the torque generated by the load
are in the same direction. In the case of pattern A,
since part 2 corresponds to the passive link, the other

1 Operating range
Bend angle θ: ±90˚
Swing angle φ: ±360˚× n (No limit)
2 The actuator and reduction gear are located in the
input-side segment to reduce the weight and sizeof moving part.
3 Reducing bearing clearance inside the joint
promotes accuracy of the motion under low
operating resistance.
4 Adoption of constant velocity joint structure allows
smooth motion in all direction for inverse input.
5 It ensures sealing capability because the parts
linking acutuator and joint are located into thejoint.

Moment (N・mm)

250

3. Analysis of Mechanism
The general-purpose mechanism analysis software"ADAMS" was used to analyze the mechanism with a
5N perpendicular load exerted at the center on the topof the output segment (Fig. 5). The support load oneach link bearings and the torque of the motor were
calculated. (Figs. 6, 7, 8)
The maximum load on the bearing was checked,
and it was found that both radial and axial loads wereapproximately 20N and the moment on the bearing
was approximately 200 N・mm. It was also found thatthe larger the bend angle θ the higher the load. (Fig.
6 shows the moment on the bearing of the passive link
(3rd linkage system) of the input-side link hub at thebend angle θ= 0˚, 45˚ and 90˚.)
The maximum torque on the motor is approximately
380 N・mm. In motor 1, the torque has the peak when

θ：0˚
θ：45˚
θ：90˚

200

150

100

50

0
0

90

180

270

360

Swing angleφ (deg)Fig. 6 Bearing load to the swing angle

300

θ：45˚
θ：90˚

Torque (N・mm)

200
100
0
0

90

180

270

360

-100
-200
-300
-400

Swing angleφ (deg)Fig. 7 The required torque of the motor 1 to the swing angle

0˚
φ
θ

300

θ：45˚
θ：90˚

Torque (N・mm)

200

Load

Passive link
3rd linkage system
Arm rotation angle(+ direction)
M2 (motor 2):
2nd linkage system

0
0

90

180

270

360

-100
-200
-300

M1 (motor 1):
1st linkage system

Fig. 5 The model of ADAMS analysis

100

-400

Swing angle φ (deg)Fig. 8 The required torque of the motor 2 to the swing angle-72-

High Angle Active Link

Output segment

Top view
Link positions generating peak torque
When the centerline of input segment bearing is at
90 degrees to the centerline of the output segment

Pattern A（φ:30˚）
Front view
Input segment

part 1

part 2part 3Motor 1

Passive link

Motor 2

Pattern B（φ:150˚）Passive link

Motor 2Motor 1

Pattern C（φ:270˚） Motor 2

Motor 1

Passive link

Perpendicular load 5N

Part 1

Part 2

Part 3Fig. 9 The link position when maximum torque is generated

two links share the torque. This causes motor 1 andmotor 2 to have the peak load at a 30-degree phasedifference. Furthermore, since the torques generated
by the load and the motor are in different directions, alarger torque is required. Therefore, each motor is
exposed to the maximum torque. The torque at thistime is about twice as large as that for patterns B andC (see Figs. 7 and 8). These must be taken into
account when designing the device.

4. Future Developments
By locating two actuators in the input segment, it
has been confirmed that the position of the output
segment can be controlled in all directions by two
degrees of freedom. We will work on refining the gear
reduction mechanism and its control method in orderto improve the accuracy and movement of the device.
The design will also be made specifically for individual
applications in hopes of further development andfuture business.

Photos of authors

Keisuke SONE

Hiroshi ISOBE

Koji YAMADANew Product Planning Department,
Research & Development Center

New Product Planning Department,
Research & Development Center

New Product Planning Department,
Research & Development Center-73-

NTN TECHNICAL REVIEW No.71（2004）[ Technical Articles ]Improvement of Leakage Magnetic Flux Resistance ofIntegrated Sensor Bearings

Takashi KOIKE*
Tomomi ISHIKAWA*
Hiroyoshi ITO**Noriyoshi MIZUTANI**

A recent trend in manufacturing is to utilize electric controls to simplify machines and ensure better reliability of thesemachines.
Because of this trend, there has been an increasing need in the marketplace for bearings with sensing functions. A
bearing with a rotation sensor that uses a magnetic encoder system is currently available. However, when using thissensor bearing near equipment that generates a magnetic flux such as a motor or magnetic clutch, the sensor may
operate incorrectly due to leakage magnetic flux.
Therefore, leakage magnetic flux needs to be minimized and mechanisms to improve leakage magnetic flux
performance are needed. We have successfully enhanced the function of the integrated sensor bearing by optimizing
the mechanical structure through magnetic field analysis. An overview of this process was introduced in NTN Technical
Review No. 69. In this paper, an electronic technique for further improvement of leakage magnetic flux resistance isintroduced.

through magnetic field analysis. An overview of thisprocess was introduced in NTN Technical Review No.
69. Due to requirements of some applications further
improvement of leakage magnetic flux resistance isrequired.
NTN has established sensor circuit technology withdrastically improved leakage magnetic flux resistancecompared to conventional bearings. An overview ofthis technology is introduced in this paper.

1. Introduction
A recent trend in manufacturing is to utilize electric
controls to simplify and ensure better reliability ofmachines and to reduce the impact on theenvironment. Because of this trend, there has been an
increasing need for bearings with sensing capabilities.
Integrated sensor bearings with speed detection
capabilities have just become available on the market.
However, when using a sensor bearing near a device
that generates a magnetic flux, such as electricmotors and magnetic clutches, the sensor may be
influenced by the magnetic field, resulting in
malfunctions such as sensor output failure. Therefore,
sensors need to have sufficient leakage magnetic flux
resistance, so that they are not influenced bymagnetic flux generated by nearby devices. NTN has
been working to enhance the leakage magnetic flux
resistance by optimizing the mechanical structure

2. Integrated Sensor Bearings
Fig. 1 shows a sectional view of an integrated
sensor bearing with leakage magnetic flux resistancecharacteristics that have been designed based on theconventional magnetic field analysis. By deploying amagnetic bypass ring that diverts leakage magnetic
flux the amount entering the Hall IC has beenminimized.

*Research & Development Center New Product Planning Department**Industrial Sales Headquarters Industrial Engineering Department-74-

Improvement of Leakage Magnetic Flux Resistance of Integrated Sensor Bearings

If the influence of an external magnetic flux is strong
enough, the alternating magnetic field on the Hall
element is affected and changes up and down (seethe red and blue lines in Fig. 2), causing the output
duty ratio to change (See the rectangular waves
indicated by red and blue lines in Fig. 2). If theexternal leakage magnetic flux increases and the sine
wave output changes considerably, loss of pulses on
the sensor output may occur.

The rotating speed detector of the bearing consistsof a magnetic encoder that has alternately magnetized
N and S poles, and a magnetic sensor. A Hall IC thatcan output a rectangular wave is normally used as themagnetic sensor. The Hall IC contains a Hall element,
amplifying circuit, Schmitt trigger circuit and output
transistor. Two types of Hall IC's are available: theunipolar type and the bipolar type. The unipolar typeturns ON/OFF according to the magnitude of themagnetic field, and the bipolar type turns ON/OFF as
the N and S poles of the magnet are alternately
activated. For a rotary sensor, the latter is often usedin order to have an output duty ratio close to 50%.
NTN uses this type, the bipolar type of Hall IC.

VH
Hall element
output voltage
VH/2

OffsetTime

0Slot on rear of sensor
Rectangularwave output

Outer ring
Magnetic sensor

TP
Tn

Fig. 2 Hall element and digital output voltage

Inner ring

4. Improvement of Leakage Magnetic
Flux Resistance

Magnetic encoder

Malfunction of the sensor output caused by external
leakage magnetic flux interference can be prevented
by canceling the sensor output that changes due tothe external leakage magnetic flux or by increasing
the magnetization intensity of the magnetic encoder.
Since increase of the magnetization intensity is limited
due to the properties of the magnetic material, electric
processing to eliminate the offset components was
implemented to improve the leakage magnetic flux
resistance.

Magnetic bypass ring

Fig. 1 Sectional view of integrated sensor bearing(Conventional design)

3. Influences of Leakage Magnetic Flux
By taking an example that uses a Hall IC as themagnetic sensor, the influences of leakage magnetic
flux on the Hall element inside the Hall IC are
explained in Fig. 2.
As the inner ring and magnetic encoder rotate,alternating magnetic fields are forced to the Hall
element, causing the output voltage to change in theform of a sine wave with a center level (reference
voltage) of VH/2, which is half of the power voltage(VH) supplied to the Hall element (See the sine wave
indicated by black line in Fig. 2). Since the output
signal of the Hall element is compared with thisreference voltage (used as the threshold level), a
rectangular wave that repeatedly turns ON and OFF is
output from the Hall IC. The output duty ratio (TP/Tn) is
approx. 50% (See the rectangular wave indicated by
the black line in Fig. 2).

4. 1 Structure
To cancel the influences of the leakage magnetic
flux on the magnetic sensor, two analog output typemagnetic sensors (e.g. Hall elements) are placed
closer to each other at different positions (at 180degree electrical angles) so that they are equally
exposed to the influences of the leakage magnetic
flux. The two sensor outputs are differential-output tocancel the offset caused by the leakage magnetic flux,
and then subjected to rectangular wave processing toprovide single-phase output. This differential output
method has doubled the output sensitivity andimproved noise resistance.

-75-

NTN TECHNICAL REVIEW No.71（2004）

Phase-A output

Electric
al
angle
180˚

Hall element

Differential amplifier
Magnetic encoder
S

N

S

N

N

Fig. 3 Schematic view of sensor

is made equal to all Hall elements, the output offset
difference between Hall elements can be reduced,
making it possible to approximate the duty ratio of therectangular output to 50%. Because of this, someideas were implemented as explained in section 4.4 toprovide an equal reference voltage for all Hall
elements.

4. 2 Circuit
Fig. 4 shows a schematic view of the detector
circuit. The structure of the circuit for one phase(phase-A) is shown. To reduce the number ofelectrical components, a comparator was used tocompare the two outputs to provide a rectangular
waveform signal. This was used instead of differential
amplification of two Hall element outputs.
Typically, an optical encoder is used separately todetect the number of revolutions. If the optical encoder
is integrated with the bearing in one unit, both bearing
and sensor will be put in the same environment which
will often hinder use of the optical encoder. Thus,sufficient care must be paid not to allow the detector
circuit to malfunction even if temperature changes in awide range. The Hall element can easily be affectedby the operating temperature, and if the ambient
temperature rises, the input resistance inside the Hall
element will drop to increase the current, resulting inlarger heat generation inside the detector. With this inmind, the Hall element drive voltage was set
according to the operating temperature range.
Two Hall element drive methods are available:
constant-current drive and constant-voltage drive. The
constant-voltage drive method was employed since
the constant-current drive method causes the Hall
output voltage and unbalanced voltage to change
excessively depending on the ambient temperature.
The Hall element's reference voltage (VH/2) varies with
the Hall element used. If the reference voltage (VH/2)

4. 3 Circuit Operation
The integrated sensor bearing with the earlier
explained built-in detector circuit was put in themagnetic flux interference test equipment 1) shown in
Fig. 5, and waveform was observed at points A, B and
C of the detector circuit (Fig. 4). The results are given
in Fig. 6. The test equipment generated a magnetic
flux by forcing current into a generating coil and theresulting magneto-motive force (AT, current×numberof coil turns) was measured when the sensor output
malfunctioned. The result was used as the criterion of
judgment for evaluation of the leakage magnetic flux
resistance. When the magneto-motive force wasincreased to expose the bearing to a magnetic flux asshown in Fig. 6 (b), it can be observed that the Hall
element's output voltage is changed upward.
However, by adopting the method that obtains
differential output of two Hall elements, a rectangular
waveform output with the same duty ratio can beobtained even if the output is exposed to the leakage
magnetic flux.

Vc

A

Rectangular wave output
VH

C

B

VH

Hall element
Vc/2

N

S

N

S

Differential amplifier (Comparator)
Rotation

Vc/2

N

S

Rotation

N

Electrical angle 180˚

Fig. 4 Schematic view of detector circuit-76-Improvement of Leakage Magnetic Flux Resistance of Integrated Sensor Bearings

Magnetic flux generating coil

Nonmagnetic section

Motor

Integrated sensor bearing

Fig. 5 Magnetic flux leakage durability test equipment

Voltage（V）

Point C

Point C

Point A

Point A

OffsetPoint B
Point B(a) Without external magnetic field (Magneto-motive force: 0AT)(b) With external magnetic field (Magneto-motive force: 400AT)

Fig. 6 Waveform of detector circuit

4. 4 Influence on Reference Voltage
700

Reference voltage (mV)The reference voltages of two Hall elements
comprising a differential circuit change according to
the leakage magnetic flux. This change was measured
and shown in Fig. 7.
From the figure, it can be understood that thereference voltages of two Hall elements change due toinfluence of leakage magnetic flux (magneto-motive
force). This change of the reference voltages is anoffset caused by the leakage magnetic flux, and thesensors interference resistance can be improved byminimizing the difference in the reference voltages ofHall elements. We have succeeded in almost
equalizing the reference voltages by adopting a circuitthat causes the reference voltage of a Hall element tomatch that of another element.

600
500
400
300
200

Hall element 1
Hall element 2

100
0
-2

-1

0

1

Magneto-motive force (kAT)

Fig. 7 Influence on reference voltage

-77-

2

NTN TECHNICAL REVIEW No.71（2004）

5. 1 Influence of Sensitivity Difference on
LeakageMagnetic Flux Resistance

5. Evaluation of Leakage Magnetic
Flux Resistance

From Fig. 8, it is understood that the leakage
magnetic flux resistance can be improved bydecreasing the sensitivity difference between two Hall
elements comprising a differential circuit. Test
condition 3 is the worst possible condition. However,under this condition, it was found that the leakage
manetic flux resistance was five times better or morethan the conventional type, and sufficient performance
was obtained without screening Hall elements. The
magneto-motive force that causes malfunction under
test condition 1 was set to 9 kAT. However, this is the
maximum magneto-motive force that the test
equipment can produce, and the new integrated
sensor bearing can withstand even higher leakage
magnetic flux. The performance can be improved up
to 30 times compared to the conventional type byminimizing the difference between Hall elements.

The Hall element's sensitivity varies with theelements used. With our method of forming a
differential circuit, the leakage magnetic flux
resistance is also influenced by the individual
difference (sensitivity difference) of two Hall elementsthat were used. Evaluation testing was conductedunder three different conditions shown in Table 1, withthe possible maximum sensitivity difference limited to20%. Fig. 8 shows the evaluation results of theleakage magnetic flux resistance of the integrated
sensor bearing equipped with a differential circuit. For
comparison, the evaluation results for conventional
bearings equipped with a magnetic bypass ring (Fig.
1) are also given.

Table 1 Sensitivity test condition
Test condition

0
10
20- (Hall IC)

Fig. 9 shows the results of the test conducted todetermine the influences of leakage magnetic flux onthe duty ratio under three conditions given in Table 1.
When the sensitivity difference between two Hall
elements is set to zero, like condition 1, the change ofthe duty ratio is small. Even in the magneto-motive
force range of ±9kAT, the change of the duty ratio iswithin ±15％. When a duty ratio of 50±15％ is set as
the allowable range, even under conditions 2 and 3,
the magneto-motive force is 1kAT or larger, which is
sufficiently improved compared to the conventional
type.

Not usedNot usedNot used
Used

9
8
7
6
5
4
3
2
1
0
1

2

3

80

Conventional
type

70

Test condition
Duty ratio (%)

Magneto-motive force
for malfunction (kAT)

1
2
3
Conventional type

5. 2 Influence of Leakage Magnetic Flux on
Duty Ratio

Hall element sensitivity
Magnetic bypass ringdifference (%)

Fig. 8 Magnetic flux leakage durability

60
50
40

Condition 1
Condition 2
Condition 3

30
20
-10

-8

-6

-4

-2

0

2

4

Magneto-motive force (kAT)

Fig. 9 Impact on duty ratio

-78-

6

8

10

Improvement of Leakage Magnetic Flux Resistance of Integrated Sensor Bearings

6. Conclusion
This paper has introduced the electrical method toimprove the leakage magnetic flux resistance ofintegrated sensor bearings. Compared to theconventional method that uses a magnetic bypass
ring, the leakage magnetic flux resistance is five times
better or more. Adoption of this method eliminates theneed for the magnetic bypass ring, and this simplifies
the structure of the bearing, resulting in easierassembly work. In addition, it is found that this method
is influenced by sensitivity difference between Hall
elements composing a differential circuit, but the duty
ratio of the sensor output does not change so much.
Since employment of this technology enables
installation of integrated sensor bearings into devices,such as motors and magnetic clutches, that generatestrong magnetic fields, the application range of
integrated sensor bearings is expected to expand.

References
1) Takashi KOIKE, Yoshitaka NAGANO, NTN Technical
Review No.69 (2001)

Photos of authors

Takashi KOIKE

Tomomi ISHIKAWA

Hiroyoshi ITONoriyoshi MIZUTANINew Product Planning Department,
Research & Development Center

New Product Planning Department,
Research & Development CenterIndustrial Engineering Department
Industrial Sales Headquarters

Industrial Engineering Department
Industrial Sales Headquarters-79-

NTN TECHNICAL REVIEW No.71（2004）[ Technical Articles ]Introduction of Grinding Swarf Recycling

Kanji NAKAMURA*

NTN has developed a system to recycle grinding swarf. Recycling has already begun.
NTN has also started a new company, UNI TOP, which will supply briquetting machines and support recycling ofbriquettes.
UNI TOP will contribute to society by reducing the environmental load of grinding swarf.

material in steel production. In addition, the coolant
separated in the process is returned to the coolant
tank to be reused again.

1. Introduction
Companies like NTN who produce precision
machinery use many grinding machines in their
manufacturing processes. This machinery produces a
lot of grinding swarf or industrial waste, most of whichis dumped into landfills. Being an ISO14001-certified
company, NTN thought that it was responsible forhelping establish recycling technology andcontributing to our recycle-based society. NTN has
been working on the development of grinding swarfrecycling technology since 1999 and has succeeded
in its practical use. In addition, not only is this systemfriendly to the environment, it also helps reduce rising
disposal costs.

3. Briquette Technology and Test
Results
Grinding swarf primarily contains grinding sludge(metallic components), coolant (oil-based or waterbased) and a small amount of grinding powder.
During the process, coolant is separated from theswarf and the metallic components are solidified into
briquettes. After completed, both the coolant and themetallic components are ready to be reused again. In
addition, the technology NTN developed produces
briquettes from grinding swarf without the use of abinder (see Fig. 2). Fig. 3 shows a magnified view ofgrinding swarf.
Fig. 4 shows a composition comparison between
grinding swarf and a briquette. The figure shows thata large amount of coolant is squeezed out of thegrinding swarf in order to create the briquettes.

2. Recycling System
Fig. 1 shows a comparison between theconventional grinding swarf dumping system and therecycling system NTN developed. Briquettes obtained
from this new system are delivered to steel
manufactures, where the briquettes are reused as raw*Engineering Department, Unitop Corporation-80-Introduction of Grinding Swarf Recycling(1) Current system
Coolant
Grinding
wheel

Coolant

Grinding swarf+ Grinding swarf

DumpingWork piece

Return of coolant
Coolant tank

Grinding machine(2) New system

Filter (filtration + press)

Truck

Original metal
Grinding swarf

Briquette

Coolant reuse
Coolant tank

Grinding machine

Briquette machineElectric furnace

Fig. 1 Comparison of current system and new system

Grinding
swarf

0.1mm

Coolant, grinding powder

Fig. 2 Briquette method

Vol %

Metal

Grinding powder

Fig. 3 Magnified view of grinding swarf

Coolant

wt %

100

100

80

80

60

60

40

40

20

20

Metal

Grinding powder

Coolant00Grinding
swarf

Grinding
swarf

Briquette

Briquette(2) Weight comparison(1) Volume comparison
Fig. 4 Comparison of compositions between grinding swarf and briquette-81-NTN TECHNICAL REVIEW No.71（2004）

4. Examples of Practical Applications
NTN has been promoting swarf recycling using thebriquette technology described earlier. Figs 5, 6 and7 show the benefits of recycling oil-based coolant
using the swarf recycling process. The examples

below are result of recycling that was started in April,
2001 on a preliminary basis and was fully
implemented in October, 2001.

Tons
25
20
15
10

Mar-02

Feb-02Jan-02

Dec-01

Nov-01

Oct-01

Sep-01

Aug-01

Jul-01

Jun-01

Apr-010

May-01

5

Fig. 5 Grinding swarf processed per month

Liters
14,00012,00010,000
8,000
6,000
4,000

Mar-02

Feb-02Jan-02

Dec-01

Nov-01

Oct-01

Sep-01

Aug-01

Jul-01

Jun-01

Apr-010

May-01

2,000

Fig. 6 Oil-based coolant reused per monthThousands of U.S.dollars
20

Fig. 7 Monthly cost savings-82-

Mar-02

Feb-02Jan-02

Dec-01

Nov-01

Oct-01

Sep-01

Aug-01

Jul-01

Jun-01

May-010

Apr-0110

Introduction of Grinding Swarf Recycling

5. Establishment of Unitop Corporation

6. Conclusion

Since grinding swarf briquette technology could be
used for various kinds of grinding swarf, it may bedifficult for machine manufacturers to assume
leadership in the practical application of thistechnology. Unlike these manufacturers, NTN
produces a large amount of grinding swarf and is in anadvantageous position to develop briquettetechnology.
To further develop this technology. NTN established
Unitop Corporation in May, 2002 with Noritake Co.,
Ltd., a manufacturer of grinding wheels, and Nicotec
Co., Ltd., a manufacturer of machines.
Briquette machines developed by Unitop have been
supplied to NTN, where they have made considerable
achievements in reducing both cost and the toll on theenvironment. In the future, NTN plans to promote fullscale sales activities in order to supply thesemachines to industrial companies.

Worldwide interest in environmental conservation is
growing. Today, companies not focused on reducing
environmental burdens will not only find theircorporate image diminished, but will also find it difficult
to manufacture their products efficiently. In thissense, the grinding swarf recycling technology that
NTN has invented and put into practical use is very
important because it can help many companies
reduce their environmental burdens on theenvironment.
To solve issues regarding the disposal of industrial
waste, NTN would like to help preserve the global
environment by providing the know-how gained
through work on grinding swarf briquette technology.

Photo of the author
Received the Minister of Economy Trade and Industry (METI)Award in the Recycling and Systems Category in 2002,
sponsored by the Ministry of Economy, Trade and Industry.
(March 11, 2003)

Kanji NAKAMURA
Engineering Department,
Unitop Corporation-83-

NTN TECHNICAL REVIEW No.71（2004）[ New Products Information ]Hub Bearing with Wireless ABS Sensor
The wireless ABS sensor module is a single unit containing a hub bearing,
hub joint, high-efficiency generator, and wireless transmitter. Use of thegenerator output as a power supply and sensor signal allows wirelesstransmission of the wheel speed signal.

Receiverunit

Features

Structure¡Compact

Generator

Hub bearing, hub joint, and sensor are integrated
into one unit.

Magnet¡Reduced assembly time
No wiring is required between the wheel and tire
housing.

¡Improved safety (Prevention of accidents
caused by damaged or broken wires)No wiring is used in moving parts.
Transmitter¡Improved design flexibility
Entire sensor is integrated into a single unit.

GEN3 hub bearing type

¡Possible to activate ABS function at very low
speeds.
It also detects wheel slippage during
acceleration.

Generator
MagnetTransmitter

GEN4 hub joint type-84-

New Products Information

GEN2 Tapered Hub Bearing for Large Commercial Vehicles
The outer ring and hub are integrated into a single lightweight greased andsealed unit with high-reliability features.

Features

Roller¡Requires no preload adjustment (Unit is designed
to have the appropriate preload after assemblyon the vehicle).
¡Ease of assembly, maintenance, and inspection.

Pulse ring

Seal
Connector

Bearing Specifications

Inner ring

¡Outer ring material: NTN's unique SC steel withexcellent fatigue strength and shock resistance¡Inner ring and roller material: NTN's unique longlife carburized steel with excellent hardness,
toughness, and shock resistance¡Lubrication: Long-life Urea grease with highfretting resistance¡Fluorocarbon rubber seals with excellent hightemperature endurance and seal lips resistant tomuddy water

Outer ringResin cage

-85-

NTN TECHNICAL REVIEW No.71（2004）

GEN3 Hub Bearing with Brake Rotor
GEN3 hub bearing and brake rotor are integrated into a single unit, resulting
in drastically reduced rotor runout.

Features

Structure¡Not as susceptible to brake shudder¡Decreases drag and reduces fuel consumption¡Eliminates the need for matching brake rotor to
hub

Applications¡Axle unit for passenger cars

-86-

New Products Information

GEN4 Hub Joint
The third generation hub bearing is integrated with the new "E Series"
constant-velocity joint, resulting in a compact, lightweight design.

Structure

Features¡High efficiency and low heat generation
Adoption of the new "E Series" constant-velocity
joint improves transmitted torque by 30% during
power conveyance. Heat generation has also
been reduced by 20˚C compared to theconventional type.
¡Smaller axial dimension
Integration of the new constant-velocity joint andbearing reduces the distance between the flange
and CVJ center by 20% or more.
¡Lightweight
Integration of the CVJ and bearing, adoption of ahollow joint shaft, and use of the new CVJ
bearing tightening method reduces the weight by
10% or more.

-87-

NTN TECHNICAL REVIEW No.71（2004）

High Efficiency Compact
Constant Velocity Joints - E SeriesThis E series constant velocity joint provides high levels of functionality andenvironmental solutions by offering reduced weight, compactness, and hightransmission efficiency.
Fixed type

EBJ（θ＝47˚）

EUJ（θ＝50˚）Plunging type

EDJ

ETJ

Comparison with NTN's conventional
CVJ product
EBJ

EUJ

EDJ

ETJ

Weight (%)

-15

-15

-10

-12

Outer diameter (%)

-7

-7

-4

-8

Temperature rise (˚C)

-20

-20

-20

―

-88-

New Products Information

PTJ (Super-Shudderless Constant Velocity Joints)
New plunging type constant velocity joints with low, constant vibration.

Features100

3rd order component of induced
cyclic axial load (N)¡Minimized vibration for drive shafts.
¡Constant induced cyclic axial load not related to

the working angle.
¡50% lower induced cyclic axial load incomparison with SFJ.
¡Identical outer diameter to SFJ

80

AC-TJ
60

FTJ
40

SFJ
20

PTJ
0

4

6

8

10

12

Working angleθ(deg)

-89-

14

NTN TECHNICAL REVIEW No.71（2004）

Super High Speed AT Spindle
Aero-static bearing spindle that can be attached to the machining
center main spindles

Features¡Can be attached to the machining center spindles (special machine tools are not needed).
¡Use of aero-static bearings deliver super-high speeds up to 150,000 min-1 with high rotational accuracy.
¡Use of a shaft-integrated shrink fit chuck enables high-accuracy and high-stiffness chucking.
¡Low vibration, low noise level, and long life

Specifications
Maximum speed 150,000 min-1
Motor
Axial ：20 N
Load capacity

Rated Power 0.6kW
Three-phase induction motor：200V AC
Water-cooled：1 to 2 L/min

Radial ：40 N

Applicable tool diameter φ1mm or smaller

Axial ：1.0 N /μm

Tool chucking method

Shaft-integrated shrink fit chuck (φ6)

Radial ：1.8 N/ μm

Supply air pressure

0.49 MPa

Static stiffness

Spindle weight

5 kg (excluding mounting shank)

Air consumption

100 L/min (A.N.R.)＊The radial load capacity and static stiffness are measured at a position 14mm from the spindle end.

-90-New Products Information

BEAREE Slide Guide
Lightweight and compact linear guide with optimal utilization of
BEAREE material.

Features

Structure¡Low friction at loads up to 50N

BEAREE UH3954¡Lighter (approx. 1/3) and less expensive than theconventional linear guide with balls¡Allows custom design of various shapes
Collar (SUS)
Table (A6063)
Rail (A6063)
BEAREE AS5000

0.25

Dynamic friction coefficient

Dynamic friction coefficient

Friction data (Reciprocating testing device)Sliding velocity：0.05mm/sec
Stroke：10mm

0.20
0.15

0.10

0.05

0.00
0

10

20

30

40

0.25

Vertical load：9.8N
Sliding velocity：150mm/sec
Stroke：120mm

0.20
0.15

0.10

0.05

0.00

50

0

Vertical load (N)

100

200

300

Cycle (×1000)

-91-

400

500