JP2000170774A - Conical roller bearing and gear shaft support device for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a conical roller bearing and a gear shaft support device for vehicle which have little torque loss by a friction and little heating, and can reduce the running-in time. SOLUTION: While the curvature radius R of the large end face 18 of a conical roller 16 is made in the value in the scope R/RBASE=0.75 to 0.87, when the distance from the top of the conical angle of the conical roller 16 to the large brim surface 13 of an inner ring 15 is made RBASE, the small brim surface 14 of the inner ring 15 is formed by a surface parallel to the small end face 19 of the conical roller 16. As a result, a torque loss owing to a sliding friction between the large brim surface 13 of the inner ring 15, and the large end face 18 of the conical roller 16, and a heating are reduced, so as to prevent the generation of a seizure. Furthermore, the running-in time when a bearing is installed is made possible to reduce.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to a tapered roller bearing and a vehicle gear shaft support device.

[0002]

2. Description of the Related Art A tapered roller bearing is a bearing suitable for applying a radial load, an axial load, and a combined load thereof. Since the load capacity is large, power transmission of a differential or a transmission in an automobile, a construction machine, or the like is performed. It is often used for gear shaft support of equipment.

FIG. 1 shows a differential of an automobile in which a gear shaft is supported by a tapered roller bearing according to an embodiment of the present invention. The differential includes a drive pinion 4 rotatably supported by two tapered roller bearings 2, 3 in a housing 1, a ring gear 5 meshing with the drive pinion 4, and a ring gear 5 attached thereto, and a pair of tapered roller bearings 6. A differential gear case 7 rotatably supported by the housing 1, a pinion 8 disposed in the differential gear case 7, and a pair of side gears 9 meshing with the pinion 8. These are housed in a housing 1 in which gear oil is sealed. This gear oil also serves as lubricating oil for the tapered roller bearings 2, 3, and 6.

FIG. 6 shows an embodiment of a conventional tapered roller bearing. The tapered roller bearing has an outer race 22 having a conical raceway surface 21 and a conical raceway surface 23.
3, an inner ring 26 provided with a large flange surface 24 on the large diameter side and a small flange surface 25 on the small diameter side, and respective raceway surfaces 21 of the outer ring 22 and the inner ring 26;
23, a plurality of tapered rollers 27 that are arranged to be able to roll freely.
And a retainer 28 that holds the tapered rollers 27 at a predetermined circumferential interval. In addition, the large collar surface 2 of the inner ring 26
The distance between 4 and the small flange surface 25 is designed to be slightly longer than the length of the tapered rollers 27.

[0005] The tapered roller 27 comprises an outer ring 22 and an inner ring 26.
Are designed to be in line contact with the respective raceway surfaces 21 and 23, and that the apex of each conical angle of the tapered roller 27 and each raceway surface 21 and 23 coincide at one point O on the center line of the tapered roller bearing. As a result, the tapered rollers 27 can roll along the respective raceway surfaces 21 and 23.

In the tapered roller bearing, each of the raceway surfaces 21, 2
Since the cone angles of No. 3 are different, the resultant of the loads applied to the tapered rollers 27 from the respective raceway surfaces 21 and 23 acts in the direction of pushing the tapered rollers 27 toward the large collar surface 24 of the inner ring 26. Therefore, the tapered roller 27 is guided by the large end surface 29 being pressed against the large flange surface 24 when the bearing is used, and the large end surface 29 and the large collar surface 2 are guided.
4 makes sliding contact.

On the other hand, since the distance between the large flange surface 24 and the small flange surface 25 of the inner ring 26 is designed to be slightly longer than the length of the tapered rollers 27, as shown in FIG. The surface 25 does not contact the small end surface 30 of the tapered roller 27, and there is a slight gap between the two. In addition, 25
Is formed as a surface inclined outward with respect to the small end surface 30 of the tapered roller 27, and the small flange surface 25 and the small end surface 30 that are not in contact with each other are not ground in the bearing manufacturing process.

When the tapered roller bearing as described above is mounted on a mounting portion, first, as shown in FIG.
6. The assembled body including the plurality of tapered rollers 27 and the retainer 28 is inserted into the raceway surface 21 of the outer race 22 from above with the large end surface 29 of the tapered rollers 27 facing upward. At this time, the plurality of tapered rollers 27 are
Because you have the freedom to
The small end face 30 comes into contact with the small flange face 25 of the inner ring 26, and the initial assembly state is established in which a gap of δ is generated between the large end face 29 and the large collar face 24 of the inner ring 26.

[0009] Next, the initial assembled state the tapered roller bearing is temporarily assembled to the mounting site of the other unit, as shown in FIG. 8 (b), while applying an axial load F _a on the end face of the inner ring 26, 50 When the running-in operation is performed at a low speed rotation of about 100 rpm, the tapered roller 27 moves axially toward the large flange surface 24 of the inner ring 26 by the gap δ, and as shown in FIG.
The large end surface 29 comes into contact with the large flange surface 24 of the inner ring 26 and settles at a regular position when the bearing is used, and a gap of δ is generated between the small end surface 30 and the small collar surface 25 of the inner ring 26. Thereafter, the tapered roller bearing is preloaded with a predetermined load in the axial direction.
This preload is performed to prevent the tapered rollers 27 from moving in the axial direction during use of the bearing, and to make the tapered rollers 27 stably come into line contact with the raceways 21 and 23 of the outer ring 22 and the inner ring 26.
The management of the preload is performed by measuring the shaft torque. When the shaft torque reaches a predetermined value, the preload operation is completed.

[0010]

The tapered roller bearing described above is used for supporting a gear shaft such as a differential rotating at high speed and high load because the large end surface of the tapered roller is in sliding contact with the large flange surface of the inner ring. The friction torque increases due to the sliding friction, and the temperature of the bearing increases due to the frictional heat, and the viscosity of the gear oil as the lubricating oil decreases, which may cause a problem due to an insufficient oil film.

In the above-described operation of mounting the tapered roller bearing on the mounting portion, the clearance between the large end surface 29 of the tapered roller 27 and the large flange surface 24 of the inner ring 26 in the initial assembly state shown in FIG. Then, there is a problem in that the running-in operation time until the tapered rollers 27 settle in the regular position shown in FIG.

Further, as shown in FIG. 7, the small collar surface 25 of the inner ring 26 is formed as a surface inclined outward with respect to the small end surface 30 of the tapered roller 27. There is also a problem that the gap between the large end surface 29 and the large flange surface 24 in the assembled state varies, and the running-in operation time until all the tapered rollers 27 settle to the proper position is further increased.

Generally, the small end surface of the tapered roller remains a forged surface, and the chamfer dimensions and shapes vary greatly. The variation in the chamfer dimension and shape exists not only between the tapered rollers but also in the circumferential direction of one tapered roller. As shown by the solid line and the dotted line in FIG. 7, if the chamfer dimensions and the shape of the small end face 30 are different, in the case of the small end face 30 shown by the solid line, the point P1 on the small end face 30 in the initial assembly state is a small flange face. Q on 25
The gap δ between the small end face 30 and the small flange face 25 when the tapered roller 27 comes into contact with one point and settles at a regular position is δ ₁ .
On the other hand, in the case of the small end surface 30 shown by the dotted line, P
2 points in contact with the point Q2, the gap [delta] is the [delta] ₂ when the tapered roller 27 has settled into the normal position. Therefore,
Depending on the chamfer dimensions and shape of the small end face 30, each tapered roller 27
Causes a variation in the time it takes for the
The running-in time becomes longer.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a tapered roller bearing and a vehicle gear shaft support device that can reduce the torque loss and heat generation due to friction and can shorten the running-in operation time.

[0015]

In order to solve the above-mentioned problems, the present invention has an outer race having a conical raceway surface, a conical raceway surface, and a large-diameter side of the raceway surface. A flange surface, an inner ring provided with a small flange surface on the small diameter side, a plurality of conical rollers arranged to be able to roll freely between a raceway surface of the outer ring and a raceway surface of the inner ring, and the conical rollers in a predetermined circumferential direction. A tapered roller bearing that includes a retainer that holds the roller at an interval, and that guides the large end surface of the tapered roller into contact with the large flange surface of the inner ring when the bearing is used. R is the radius of curvature of the large end face of the tapered roller, and R _{BASE is} the distance from the apex of the cone angle of the tapered roller to the inner ring large collar surface.
Then, the configuration in which R / R _BASE is in the range of 0.75 to 0.87 is adopted.

The reason why the small collar surface of the inner ring is formed as a surface parallel to the small end surface of the tapered roller is as follows. FIG. 2 (b)
As shown in the enlarged view, the small collar surface 14 of the inner ring 15 is made to be a surface parallel to the small end surface 19 of the tapered rollers 16 arranged on the raceway surface 12, so that the tapered rollers 16 in the above-mentioned initial assembly state are formed. Chamfer of the tapered roller 16 small end face 19 with respect to the gap between the large end face 18 and the inner ring 15 large flange face 13 (equal to the gap δ between the small end face 19 and the inner ring 15 small collar face 14 when the tapered roller 16 is settled at a regular position). The influence of variations in size and shape can be eliminated. That is, as shown by the dotted lines in FIG. 2B, even if the chamfer dimensions and shapes of the small end faces 19 are different, the small end faces 19 parallel to each other are initially set.
And the small collar surface 14 are in surface contact, and the large end surface 1
8 and the large flange surface 13 are always constant, and each tapered roller 1
It is possible to eliminate the variation in the time required for 6 to settle in the regular position and shorten the running-in time.

The ratio R / R _BASE between the radius of curvature R of the large end face of the tapered roller and the distance R _BASE from the vertex of the cone angle of the tapered roller to the inner ring large flange surface is in the range of 0.75 to 0.87. This is for the following reasons.

FIG. 4 shows the result of calculating the oil film thickness t formed between the inner ring large flange surface and the tapered roller large end surface using Karna's formula. The vertical axis indicates the ratio t / t ₀ to the oil film thickness t ₀ when R / R _BASE = 0.76. The oil film thickness t becomes maximum when R / R _BASE = 0.76, and R / R
_{When BASE} exceeds 0.9, it decreases sharply.

FIG. 5 shows the result of calculating the maximum Hertz stress p between the inner ring large flange surface and the tapered roller large end surface. The vertical axis is shown in FIG.
Similarly to the above, it is indicated by the ratio p / p ₀ to the maximum Hertz stress p ₀ when R / R _BASE = 0.76. Maximum Hertz stress p
Decreases monotonically with increasing R / R _BASE .

In order to reduce torque loss and heat generation due to sliding friction between the inner ring large flange surface and the tapered roller large end surface, it is desirable to increase the oil film thickness t and reduce the maximum Hertz stress p. The present inventors referred to the calculation results of FIGS. 4 and 5, and based on the seizure resistance test results shown in Table 1 below,
The appropriate range of R / R _BASE was determined to be 0.75 to 0.87. In the conventional tapered roller bearing, the value of R / R _BASE is designed in the range of 0.90 to 0.97.

[0021] The surface roughness R _a of the inner ring large rib surface 0.05
The thickness of the oil film between the inner ring large flange surface and the tapered roller large end surface t
Therefore, the lubricating state between these surfaces can be maintained in an appropriate state.

The reason for setting the surface roughness _Ra to 0.05 μm or more is as follows. As shown in FIG. 8 (b), at the time of installation of tapered roller bearings, while applying an axial load F _a on the end face of the inner ring 26, 50～100Rp
The running-in operation is performed at a low speed of about m. Surface roughness R
_{If a} is less than 0.05 μm,
Since the lubrication state between the large flange surface 24 of the inner ring 26 and the large end surface 29 of the tapered roller 27 is a mixed lubrication of fluid lubrication and boundary lubrication, the friction coefficient greatly fluctuates, and the variation in the measured shaft torque is reduced. And the precision of the preload control deteriorates. When _Ra is 0.05 μm or more, the lubrication state becomes boundary lubrication, the friction coefficient is stabilized, and accurate preload management can be performed. At a rotational speed under normal bearing operating conditions exceeding 100 rpm, the large flange surface 24 and the large end surface 2
Since a sufficient oil film is formed between the surfaces 9, the lubrication state between these two surfaces is fluid lubrication, and the friction coefficient is reduced.

[0023] was not more than 0.20 [mu] m the surface roughness R _a, when R _a is more than 0.20 [mu] m, when the bearing unit is increased temperatures, lubricating oil was reduced viscosity at a high speed rotation region,
This is because the oil film thickness t becomes insufficient and seizure easily occurs.

The gap δ between the small flange surface of the inner ring and the small end surface of the tapered roller, which is formed when the large end surface of the tapered roller contacts the large flange surface of the inner ring, is δ ≦ 0. By restricting the dimension within the range of 4 mm (claim 3), the number of rotations required until the tapered rollers in the running-in operation settle in a proper position can be reduced, and the running-in time can be shortened. The allowable maximum value of 0.4 mm for the gap δ is determined based on the result of a running-in test described later.

By making the small collar surface of the inner ring a grinding surface or a turning surface (Claim 4), the gap δ between the small collar surface of the inner ring and the small end surface of the tapered roller can be accurately controlled. can do.

Further, the present invention is a vehicle gear shaft support device in which a gear shaft is rotatably supported by a tapered roller bearing in a housing in which gear oil is sealed, wherein the tapered roller bearing has a conical shape. An outer ring having a raceway surface, an inner ring having a conical raceway surface, a large flange surface on the large diameter side and a small flange surface on the small diameter side of the raceway surface, a raceway surface of the outer ring and a raceway surface of the inner ring A plurality of conical rollers arranged to be able to roll freely between
A retainer for holding the tapered rollers at a predetermined circumferential interval, wherein the large end surface of the tapered rollers is brought into contact with the large flange surface of the inner ring when the bearing is used, and the small flange surface of the inner ring is When formed from a surface parallel to the small end surface of the tapered roller, the radius of curvature of the large end surface of the tapered roller is R, and the distance from the apex of the cone angle of the tapered roller to the inner ring large flange surface is R _BASE ,
By adopting a configuration in which the R / R _{BA SE} in the range of from 0.75 to 0.87 (claim 5), it is possible to improve the durability by reducing the heat generation and torque loss due to friction of the gear shaft supporting device, The running-in time can be reduced.

[0027] In the gear shaft support system the vehicles, 0.05～0.20Myuemu a surface roughness R _a of the large rib surface of the inner ring
(Claim 6), the lubrication state between the inner ring large flange surface and the tapered roller large end surface can be maintained in an appropriate state, and the durability of the gear shaft support device can be further increased.

In the vehicle gear shaft supporting device, the large end surface of the tapered roller may be formed when the large end surface of the tapered roller comes into contact with the large flange surface of the inner ring. By regulating the gap δ between δ to 0.4 mm (claim 7), the number of rotations required for the tapered rollers in the running-in operation to settle in a proper position is reduced, and the running-in time is reduced. Can be shortened.

In the vehicle gear shaft support device, the small collar surface of the inner ring may be a grinding surface or a turning surface (Claim 8), so that the small collar surface of the inner ring and the small end surface of the tapered roller may be formed. Can be accurately managed.

[0030]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows a differential of an automobile, as described above, and the drive pinion 4.
The gear shaft supporting device using the tapered roller bearings 2, 3, and 6 of the embodiment is adopted to support the shaft of the differential gear case 7 to which the ring gear 5 is attached, and the shaft of the differential gear case 7.

FIG. 2A shows a tapered roller bearing 6 as a typical example. The tapered roller bearing 6 has a conical raceway surface 10.
And a conical raceway surface 12 having a large flange surface 13 on the large diameter side and a small flange surface 14 on the small diameter side.
Are provided, a plurality of tapered rollers 16 arranged to be able to roll freely between the raceways 10, 12 of the outer ring 11 and the inner ring 15, and the tapered rollers 16 are held at predetermined circumferential intervals. And a retainer 17.

The small collar surface 14 of the inner ring 15 is shown in FIG.
As shown in an enlarged manner, the grinding surface is finished in parallel with the small end surface 19 of the tapered rollers 16 arranged on the raceway surface 12,
In the initial assembly state shown by the chain line in the drawing, the small end face 19 of the tapered roller 16 is in surface contact with the small end face 19 of the tapered roller 16, and the gap δ between the tapered roller 16 and the small end face 19 in the state where the tapered roller 16 shown by the solid line is settled at a regular position is formed. , Δ ≦ 0.4 mm.

As shown in FIG. 3, the tapered rollers 16
The vertices of the cone angles of the raceway surfaces 10 and 12 of the outer ring 11 and the inner ring 15 coincide at one point O on the center line of the tapered roller bearing 6, and the radius of curvature R of the large end surface 18 of the tapered roller 16 and the point O to the inner ring 15 ratio R / R to the distance R _BASE to the large brim 13
_BASE is manufactured in the range of 0.75 to 0.87. The large flange surface 13 is ground to _a surface roughness Ra of 0.12 μm.

Although not shown, the tapered roller bearings 2 and 3 are manufactured with the same configuration and specifications.

In this embodiment, the small flange surface of the inner ring is a ground surface, but it may be a turned surface for cost reduction. The present invention can be applied to various types of tapered roller bearings.

Examples and comparative examples will be described below.

[0037]

EXAMPLES 2 and 3, the radius of curvature R of the large end faces of tapered rollers, wherein R / R _BASE = from 0.75 to .87
It enters the range of the surface roughness R _a of the large rib surface of the inner ring is 0.12
The tapered roller bearing in which the small flange surface is formed by a grinding surface parallel to the small end surface of the tapered roller, and the gap δ is within the dimension regulation range of δ ≦ 0.4 mm (Example in Table 1) 1 to
4) was prepared. Bearing dimensions are 40m inside diameter
m, outer diameter 68 mm.

[0038]

[Comparative Example] The value of R / R _BASE is out of the range of the present application, and the small collar surface of the inner ring is inclined outward with respect to the small end surface of the tapered roller, and the gap δ exceeds 0.4 mm. (Comparative Examples 1 to 3 in Table 1) were prepared. The dimensions of each bearing are the same as in the embodiment.

An anti-seizure test was performed on the tapered roller bearings of the above Examples and Comparative Examples using a rotation tester.
For the tapered roller bearings of Example 2 and Comparative Example 2,
A running-in test was also performed. The number of samples in the running-in test was 66 for Example 2 and 1 for Comparative Example 2.
The number was set to 0. The test conditions for the seizure resistance test are as follows. Load: 19.61 kN Revolution: 1000-3500 rpm Lubricating oil: Turbine VG56 (lubricating amount 40 ml / min, lubricating temperature 40 ° C. ± 3 ° C.)

[0040]

[Table 1]

Table 1 shows the test results. The seizure in the anti-seizure test occurred between the large collar surface of the inner ring and the large end surface of the tapered rollers.

In each of the tapered roller bearings of the examples, the critical rotation speed at which seizure occurs in the seizure resistance test is 2700 rpm.
m, which indicates that the frictional resistance between the inner ring large flange surface and the tapered roller large end surface is small. On the other hand, in the tapered roller bearing of the comparative example, the critical rotation speed at which seizure occurs is 2500 rpm or less, which may cause a problem under normal use conditions such as differential. Surface roughness R _a rough Comparative Example 3 of the large rib surface shows a low Seizure limit speed than Comparative Example 2 having the same radius of curvature R.

The result of the running-in test is that, in the comparative example, the average value of the number of rotations until the tapered roller settles to the normal position is 6 times, whereas in the example, about half of the number of rotations is 2.
96 times. In the example, the standard deviation of the variation in the number of rotations is also small, and it can be seen that the running-in time can be stably reduced.

[0044]

As described above, in the tapered roller bearing of the present invention, the radius of curvature R of the large end face of the tapered roller is set to the value of R / R _BASE.
= 0.75 to 0.87, and the inner ring small flange surface is formed by a surface parallel to the tapered roller small end surface, so that the sliding friction between the inner ring large flange surface and the tapered roller large end surface. To prevent seizure by reducing torque loss and heat generation.
In addition, the running-in time can be shortened, and the work of mounting the bearing can be made more efficient. Further, the durability of the vehicle gear shaft support device can be improved.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional view of a differential into which a gear shaft support device according to an embodiment is incorporated.

2A is a longitudinal sectional view showing the tapered roller bearing of FIG. 1, and FIG. 2B is an enlarged sectional view of a main part of a.

FIG. 3 is a sectional view for explaining the design specifications of the tapered roller bearing of FIG. 1;

FIG. 4 is a graph showing a relationship between a radius of curvature of a tapered roller large end face and an oil film thickness.

FIG. 5 is a graph showing the relationship between the radius of curvature of the tapered roller large end face and the maximum Hertz stress;

FIG. 6 is a partially omitted longitudinal sectional view showing a conventional tapered roller bearing.

FIG. 7 is an enlarged sectional view of a main part of FIG. 6;

FIGS. 8A, 8B, and 8C are cross-sectional views each illustrating a mounting operation of a tapered roller bearing.

[Explanation of symbols]

 Reference Signs List 1 housing 2, 3 tapered roller bearing 4 drive pinion 5 ring gear 6 tapered roller bearing 7 differential gear case 8 pinion 9 side gear 10 raceway surface 11 outer ring 12 raceway surface 13 large flange surface 14 small flange surface 15 inner ring 16 tapered roller 17 retainer 18 Large end face 19 Small end face 21 Track face 22 Outer ring 23 Track face 24 Large collar face 25 Small collar face 26 Inner ring 27 Tapered roller 28 Cage 29 Large end face 30 Small end face

 ──────────────────────────────────────────────────続 き Continued on the front page F-term (reference) 3D039 AB01 AC22 AC86 3J101 AA16 AA25 AA32 AA42 AA54 AA62 BA05 BA53 BA57 DA11 FA33 FA46 GA01 GA51

Claims (8)

[Claims] 1. An outer ring having a conical raceway surface, an inner race having a conical raceway surface, a large flange surface provided on a large diameter side of the raceway surface, and a small collar surface provided on a small diameter side of the raceway surface, and an outer ring A plurality of tapered rollers arranged to be able to roll freely between the raceway surface of the inner ring and the raceway surface of the inner ring,
A retainer for holding the tapered rollers at a predetermined circumferential interval, wherein the large end surface of the tapered roller contacts and is guided by the large flange surface of the inner ring when the bearing is used. The surface is formed by a surface parallel to the small end surface of the tapered roller, the radius of curvature of the large end surface of the tapered roller is R, and the distance from the vertex of the cone angle of the tapered roller to the inner ring large flange surface is R _BASE . A tapered roller bearing wherein R / R _BASE is in the range of 0.75 to 0.87. 2. A surface roughness Ra of _a large collar surface of the inner ring is 0.
2. The tapered roller bearing according to claim 1, wherein the tapered roller bearing is formed in a range of 0.5 to 0.20 [mu] m. 3. A gap δ between the small flange surface of the inner ring and the small end surface of the tapered roller, formed when the large end surface of the tapered roller contacts the large flange surface of the inner ring, is δ ≦. 0.4mm
3. The tapered roller bearing according to claim 1, wherein the tapered roller bearing is restricted to the dimension range described in claim 3. 4. The tapered roller bearing according to claim 1, wherein the small collar surface of the inner ring is a grinding surface or a turning surface. 5. A vehicle gear shaft supporting device in which a gear shaft is rotatably supported by a tapered roller bearing in a housing in which gear oil is sealed, wherein the tapered roller bearing comprises:
An outer ring having a conical raceway surface, an inner ring having a conical raceway surface, a large flange surface on a large diameter side and a small flange surface on a small diameter side of the raceway surface, and a raceway surface and an inner ring of the outer ring A plurality of tapered rollers that are arranged to be able to roll freely between the raceway surfaces, and a retainer that holds the tapered rollers at predetermined intervals in the circumferential direction. In those guided in contact with the large collar surface, the small collar surface of the inner ring is formed by a surface parallel to the small end surface of the tapered roller, and the radius of curvature of the large end surface of the tapered roller is R, When the distance from the apex of the cone angle to the inner ring large collar surface is R _BASE , R / R _{BA SE}
In the range of 0.75 to 0.87. 6. The method of Claim 5, wherein the inner ring of the large rib surface of the surface roughness R _a vehicle gear shaft supporting apparatus which is formed in a range of 0.05～0.20Myuemu. 7. The tapered roller according to claim 5, wherein a large end surface of the tapered roller is formed when the large end surface of the tapered roller comes into contact with a large flange surface of the inner ring. A gear shaft support device for a vehicle, in which a gap δ between them is regulated within a dimension range of δ ≦ 0.4 mm. 8. The vehicle gear shaft supporting device according to claim 5, wherein the small collar surface of the inner ring is a grinding surface or a turning surface.