JPH07238926A - Ball bearing device - Google Patents

Abstract

PURPOSE:To improve durability by supporting an axial load by the first ball bearing which is larger than a second ball bearing and whose contact angle direction is different from the second ball bearing, providing plus space by face assembly in a bearing, and specifying the outer wheel of the second ball bearing, the cross sectional curvature radius of an inner wheel track, and ball outer diameter. CONSTITUTION:A ball bearing device is provided with first and second ball bearings 5a, 6a of an angular, which is provided between a shaft 2 outer circumferential surface and a housing 3 inner circumferential surface, and whose contact angle directions are different from each other, and an axial load applied from an outside to constant direction is supported between the shaft 2 and the housing 3 at the time of using by the first ball bearing 5a having the contact angle larger than that of the second ball bearing 6a. Relation of the cross sectional curvature radius re2 of the outer wheel track 14 of the second ball bearing 6a, the cross sectional curvature radius ri2 of an inner wheel track 15, and the outer diameter (d) of a ball 12 is set to 0.53d<=ref<=0.56d, 0.505d<=ri2<=0.52d. It is thus possible to ensure the load capacity of the first ball bearing 5a, an reduce axial direction component force applied on the second ball bearing 6a. Therefore, it is possible to reduce contact surface pressure between the balls 11, 12 of the ball bearings 5a, 6a and the outer wheel and inner wheel tracks 14, 15 so as to prevent reduction of the life.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION The ball bearing device according to the present invention is used to rotatably support a rotary shaft of a screw compressor.

[0002]

2. Description of the Related Art A ball bearing device as shown in FIG. 1 has been conventionally used to support a shaft that rotates at a high speed, such as a rotary shaft of a screw compressor. This ball bearing device is provided between the outer peripheral surface of a rotating shaft 2 to which a rotor 1 constituting a screw compressor is fixed and the inner peripheral surface of a housing 3. Incidentally, reference numeral 4 in FIG. 1 denotes a roller bearing for supporting the radial load F _r .

The ball bearing device which is the object of the present invention is for supporting an axial load F _a in the axial direction of the rotary shaft 2 (left and right direction in FIG. 1), and each is an angular type first. , The second ball bearings 5, 6 are combined. Contact angle α between the first and second ball bearings 5 and 6
The directions of (see FIG. 2 to be described later) are opposite to each other, and front combinations (hereinafter, referred to as “D
F ”. ). Therefore, when the rotary shaft 2 is to be displaced to the left in FIG. 1, when the first ball bearing 5 on the left side in FIG. , The second ball bearing 6 on the right side supports the axial load to prevent the rotary shaft 2 and the rotor 1 from being displaced with respect to the housing 3.

As a combination type which can support an axial load in any direction by combining a pair of angular type ball bearings, contrary to the structure shown in FIG. There is a so-called back surface combination in which the back surfaces of the ball bearings are combined so as to face each other. However, when such a back surface combination structure is adopted, the interval between the action points on the inner ring side of the first and second ball bearings becomes large, and the bending rigidity with respect to the inclination of the rotating shaft becomes large. . Therefore, when the back surface combination is used to support a member having a relatively large inclination, such as the rotary shaft 2 of the screw compressor, an excessive surface pressure tends to act on the action point portion. Excessive surface pressure is not preferable because it causes abnormal heat generation and shortens the fatigue life of the ball rolling surface and raceway surface. Therefore, the ball bearing device to which the present invention is applied is a DF as shown in FIG.

The ball bearing device to which the present invention is applied has the rolling surfaces of the balls 11 and 12 constituting the first and second ball bearings 5 and 6 and the inner ring raceway 15 of the inner ring 10 and the outer peripheral surface. , 15, and the outer races 13 and 13 and the outer raceways 14 and 14 on the inner peripheral surfaces thereof, a positive clearance (actually existing clearance) is left. The reason for this is to prevent heat generation and reduction of the peeling life. The screw compressor in which the ball bearing device which is the subject of the present invention is incorporated rotates at an extremely high speed (for example, d _m n is 700,000 to 2,000,000 mm · rpm). Therefore, when preload is applied to each of the balls 11 and 12 and the first and second ball bearings 5 and 6 having a so-called negative gap are used, these first and second ball bearings are pre-loaded. The internal load of the ball bearings 5 and 6 increases, which also causes abnormal heat generation and shortened fatigue life.

Therefore, in the ball bearing device of the present invention, the angular type first and second ball bearings 5 and 6 each having a positive gap are used in the DF. The rotary shaft 2 supported by this ball bearing device has an axial load F _{a in a} substantially constant direction (from right to left in FIG. 1) accompanying the rotation of the rotor 1 when the screw compressor is used. Is added. This is because the direction of pressure applied to the rotor 1 is fixed. In FIG. 1, 7 and 8 are spacers, and 9 is a presser foot.

By the way, as described above, in the ball bearing device formed by combining the first and second ball bearings 5 and 6 each of which is an angular type with DF, when the rotary shaft 2 rotates at a high speed, It is not always possible to obtain a sufficient bearing life. Regarding the cause of the shortening of bearing life with high-speed rotation,
This will be described with reference to FIG.

FIG. 2 is a diagram for explaining the forces applied to the first and second ball bearings 5 and 6 based on the centrifugal force generated at high speed rotation. When the rotary shaft 2 (FIG. 1) is rotated at a high speed, the balls 11 and 1 that form the first and second ball bearings 5 and 6
A centrifugal force F _c1 is applied to 2. And, the balls 11 and 12 constituting these first and second ball bearings 5 and 6 have the centrifugal force F.
The force of Q _in1 is applied to the outer ring raceways 14, 14 based on _c1 from the contact angle α direction. And these first,
The balls 11 and 12 which form the second ball bearings 5 and 6 are each outer ring raceway 1 by the component force F _ain1 of this force Q _{in1 in} the axial direction.
It is pressed against 4, 4 in the axial direction.

Due to this component force, an internal axial load corresponding to F _ain1 + F _ain1 is generated in the first and second ball bearings 5 and 6 which are combined in series with each other, and the outer ring raceways 14 and 14 are formed. The contact surface pressure with the balls 11 and 12 is increased, and the fatigue life of these first and second ball bearings 5 and 6 is reduced. The calculation of the fatigue life of the ball bearing considering the centrifugal force of the ball at high speed is described in "The Life o", which was published in Trans. ASME published in July 1952.
f High-Speed Ball Bearings ”.

Balls 11 and 12 constituting the first and second ball bearings 5 and 6 in order to prevent a decrease in life due to the above-mentioned causes.
Made of lightweight ceramic material, or balls 11,
Attempts have also been made to reduce the centrifugal force F _c1 by using a small diameter as 12. Furthermore, attempts have been made to reduce the component force F _{ain1 in the} axial direction based on the centrifugal force F _c1 by reducing the contact angle α between the first and second ball bearings 5 and 6. .

[0011] However, the balls 11 and 12, when made by a large ceramic material modulus, in the first ball bearing 5 side in particular supporting the external loads (e.g. the axial load F _a), the outer ring raceway The contact surface pressure between 14 and the ball 11 is
Increased compared to using balls made of bearing steel. As a result, the fatigue life of the first ball bearing 5 is reduced, and the fatigue life of the entire ball bearing device formed by combining the first and second ball bearings 5 and 6 is reduced. Further, if balls 11 and 12 having a small diameter are used or the contact angle α is reduced, the load capacity in the axial direction of the first and second ball bearings 5 and 6 (basic dynamic load rating) becomes small, which also causes fatigue. The life will be shortened.

In order to reduce the internal axial load due to such centrifugal force and prolong the fatigue life of the first and second ball bearings 5 and 6, there are disclosed in Japanese Unexamined Patent Publication Nos. 5-248431 and 5-248431.
Japanese Patent No. 280482 describes a ball bearing device as shown in FIGS. In the case of the ball bearing apparatus is provided in axial load F _a the bears load applied from the outside to the rotary shaft 2 (left side in FIGS. 3-4) at the time of use, the contact angle of the first ball bearing 5a alpha than _1, the axial load F anti-load side without supporting the _a small contact angle alpha ₂ of the second ball bearing 6a provided (right side in FIGS. 3-4) at the time of use (alpha
₂ <α ₁ ).

As described above, the contact angle α of the first ball bearing 5a
_By setting ₁ to be larger than the contact angle α ₂ of the second ball bearing 6a, the load capacity of the first ball bearing 5a can be sufficiently secured. In this case, the contact angle α ₁ is the same as the contact angle α of the first and second ball bearings 5 and 6 incorporated in the ball bearing device generally known from the prior art (α = α). ₁ ), the axial component force acting on the first ball bearing 5a based on the centrifugal force is F _ain1 . The axial component force applied to the second ball bearing 6a is F _ain2 . In this case, the contact angle α _{2 of} this second ball bearing 6a
Is smaller than the contact angle α _{1 of} the first ball bearing 5a.
Therefore, the axial component force F _ain2 applied to the second ball bearing 6a is smaller than the axial _component force F _ain1 acting on the second ball bearing 6 of the conventional device (F _ain2 <F
_ain1 ) As a result, the first and second ball bearings 5a, 6a
The internal axial load due to the centrifugal force generated in the ball bearing device formed by combining the above is F _ain1 + F _ain2 , and the internal axial load F _ain1 + F _ain1 generated in the conventional device is generated.
Will be smaller than.

For this reason, the first and second ball bearings 5a and 6 are provided.
The contact surface pressure between the balls 11 and 12 forming a and the outer ring raceways 14 and 14 decreases. Then, the first and second ball bearings 5a,
The rolling fatigue life of 6a is suppressed from being shortened.
The life of the ball bearing device formed by combining the second ball bearings 5a and 6a can be prevented from being shortened. Although the load capacity in the axial direction of the second ball bearing 6a having a small contact angle α ₂ is small, the second ball bearing 6a has a function as a kind of backup bearing that bears a rare light load. Since it is not, the reduction of load capacity is not a problem.

[0015]

However, in the case of the ball bearing device described in each of the above publications, the contact angle α ₂ of the second ball bearing 6a on the anti-load side is reduced, The durability of the second ball bearing 6a tends to be insufficient. The reason why the durability of the second ball bearing 6a having a small contact angle α ₂ is insufficient is as follows.

As described above, the first and second ball bearings 5a, 6a constituting the ball bearing device of the present invention have a positive clearance. Then, during operation of the screw compressor, almost no axial load is applied to the second ball bearing 6a on the non-load side (the radial load is also supported by the roller bearing 4 so that it is hardly applied). Therefore, the balls 12 that compose the second ball bearing 6a have a centrifugal force F _c2 (FIG. 4).
5, the balls 12 tend to be displaced outward in the diametrical direction, and as shown in FIG. 5, each ball 12 moves to the large diameter side end portion (the left end portion in FIG. 5) of the outer ring raceway 14 on the inner peripheral surface of the outer ring 13. Further, in this state, the rotation shaft 2 (see FIG. 3) to the first ball bearing 5
Due to the axial load F _a (FIG. 4) applied to the inner ring 10 a, the inner ring raceway 15 on the outer peripheral surface of the inner ring 10 tends to move away from the balls 12.

As a result, as shown in FIG.
The contact angle between the outer ring raceway 14 and the outer ring raceway becomes 0 degrees, and each ball 1
The second rolling surface and the inner ring raceway 15 tend to separate from each other.
In this state, since each ball 12 is hard to roll, each ball 1
Slip occurs at the contact portion between the rolling surface 2 and the inner ring raceway 15. Then, the contact portion is abraded by the slip, and the second ball bearing 6a is damaged early. Therefore, when a structure in which the contact angle α ₂ of the second ball bearing 6a on the anti-load side is reduced as shown in FIGS. 3 to 4 is adopted, in order to prevent early damage due to the slip, The second ball bearing 6a and the load-side first ball bearing 5a combined with the second ball bearing 6a must be used in a state in which a preload is applied (a state in which a minus clearance is provided). It was

However, as described above, the first and second ball bearings 5 are
If the rotating shaft 2 of the screw compressor is supported in a state in which a preload is applied to a and 6a, the internal load of these first and second ball bearings 5a and 6a increases based on the preload, resulting in abnormal heat generation and fatigue life. The cause of the decrease is as described above. The ball bearing device of the present invention has been invented in view of the above circumstances.

The above-mentioned paper "The Life of High-Speed"
The theory presented in "Ball Bearings" is the basis for obtaining the calculated life described in this specification. However, this paper does not disclose the theory of the case where the ball bearing is operated without a preload and with a positive clearance. Further, there is no description of a method of reducing the influence of centrifugal force by changing the conditions such as contact angles of a plurality of combined ball bearings, and, of course, neither the interests of such a method nor its solution are described.

Further, as a structure in which a plurality of ball bearings having different contact angles, ball diameters, etc. are combined, in addition to the respective publications mentioned above, "PUMPAC The MRC B" issued by MRC Bearing Services
earing System ”, 1985 FAG Catalog 02105 /
2EA, page 93, Japanese Utility Model Publication No. 53-97701, John
Rolling Bearing Analysi, published by Wiley & Sons, Inc.,
s ”by Eiichi Watabayashi,“ How to choose and use rolling bearings ”
Those described in, etc. are conventionally known. However, the structures described in each of these publications have different uses from the ball bearing device of the present invention. Therefore, there is no description in any of these publications about the above-mentioned wear due to slip and its solution.

[0021]

All of the ball bearing devices of the present invention are provided between the outer peripheral surface of the shaft and the inner peripheral surface of the housing in the same manner as the conventional ball bearing devices described in the above-mentioned respective publications. It is equipped with angular type first and second ball bearings having different contact angles from each other, and the axial load applied from the outside in a substantially constant direction between the shaft and the housing during use is The bearing is supported by a first ball bearing having a contact angle larger than that of the ball bearing.

Particularly, in the ball bearing device of the present invention, the first and second ball bearings are arranged in a DF and each have a positive clearance. When the radius of curvature of the cross-sectional shape of the outer ring raceway of the second ball bearing is r _e2 and the outer diameter of the balls forming the second ball bearing is d, 0.53 d ≦ r _e2 ≦ 0 0.55d ≦ r, where r _i2 is the radius of curvature of the cross-sectional shape of the inner ring raceway of the second ball bearing, and d is the outer diameter of the balls that make up the second ball bearing. It is characterized by satisfying _i2 ≤ 0.52d.

[0023]

The ball bearing device of the present invention configured as described above is
Similar to the ball bearing device described in each of the above publications, the contact angle of the first ball bearing that bears an axial load applied from the outside during use is large, and the load capacity of this first ball bearing is large, so it is sufficient. Can support large axial loads.
In addition, the second ball bearing that does not support this axial load during use suppresses an increase in the internal axial load due to centrifugal force. Therefore, it is possible to prevent the fatigue life of the ball bearing device as a whole by combining a pair of angular type ball bearings from decreasing.

Particularly, in the case of the ball bearing device of the present invention, since the first and second ball bearings are arranged in the DF and have a positive gap, the ball bearing device is also operated. No preload is applied, and heat generation and fatigue life reduction due to preload do not occur. Moreover, since the radii of curvature of the outer ring raceway and the inner ring raceway of the second ball bearing are regulated, slippage does not occur between the ball and the raceway of the second ball bearing to which little load is applied during use.

That is, since the radius of curvature of the cross-sectional shape of the outer ring raceway of the second ball bearing is increased, even if the ball is displaced radially outward due to centrifugal force, this ball and the outer ring raceway The contact angle of is not 0 degree. In other words, the ball and the outer ring raceway are always in contact with each other with a contact angle. Therefore,
The rolling surface of the ball is kept in contact with the outer ring raceway and the inner ring raceway at all times, and the ball surely rolls when the second ball bearing rotates. As a result, abnormal slippage does not occur between the rolling surface and the outer ring raceway and the inner ring raceway, and abnormal wear due to this slippage can be prevented.

[0026]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS The ball bearing device of the present invention, like the ball bearing devices described in the above-mentioned respective publications, bears an axial load F _a applied to the rotary shaft 2 from the outside during use, as shown in FIGS. to, than the contact angle alpha ₁ of the first ball bearing 5a provided on the load side (left side in FIGS. 3-4), no support the axial load F _a at the time of use, the anti-load side (in FIGS. 3-4 The contact angle α ₂ of the second ball bearing 6a provided on the right side is made small (α
₂ <α ₁ ).

Thus, the contact angle α _{1 of} the first ball bearing 5a is
Is larger than the contact angle α ₂ of the second ball bearing 6a, the load capacity of the first ball bearing 5a can be sufficiently secured. Since the contact angle α ₂ of the second ball bearing 6a is smaller than the contact angle α _{1 of} the first ball bearing 5a, the axial _component force F _ain2 applied to the second ball bearing 6a is Axial component force F acting on the second ball bearing 6 (FIGS. 1-2) of the conventional device
_It is smaller than _ain1 (F _ain2 <F _ain1 ). As a result,
The internal axial load due to the centrifugal force generated in the ball bearing device formed by combining the first and second ball bearings 5a and 6a is smaller than that in the conventional device shown in FIGS.

For this reason, the contact surface pressure between the balls 11 and 12 forming the first ball bearings 5a and 6a and the outer ring raceways 14 and 14 and the inner ring raceways 15 and 15 decreases. And first,
It is possible to prevent the rolling fatigue life of the second ball bearings 5a and 6a from being shortened, and to suppress the life reduction of the ball bearing device formed by combining the first and second ball bearings 5a and 6a, each of which is an angular type. You can

For example, the inventor of the present invention has an inner diameter of 25 mm and an outer diameter of 62
2 mm-width, 17 mm-width angular contact ball bearings (Type 7305)
First, in the case where it is mounted on a rotary shaft that rotates at a rotation speed of 23000 rpm while receiving an axial load of 67 kgf so that the measured axial gap during manufacture will be 0.030 mm Table 1 below shows the results of calculating the rolling fatigue life of the second ball bearing.
The contact angles of the first ball bearings (load-side bearings) that receive an axial load are all 30 degrees, and the contact angles of the second ball bearings (anti-load-side bearings) are 15, 30, and 40 degrees. Calculated for three types.

During this calculation, the tightening margin between the outer peripheral surface of the rotary shaft 2 and the inner peripheral surfaces of the inner rings 10, 10 is 0.012 mm, and the inner peripheral surface of the housing 3 and the outer peripheral surfaces of the outer rings 13, 13 are the same. Between the outer ring 13, the temperature of 80 ℃, the inner ring 1
The temperature of 0 and 10 was 85 ° C., the material of the rotary shaft 2 was steel, and the material of the housing 3 was cast iron. Note that d _m n (product of pitch circle diameter d _m (mm) of balls 11 and 12 and rotation speed n (rpm))
Is about 1 million. Also, the radial load F _r is 0
And These conditions are general conditions as a ball bearing device for a screw compressor. Then, according to such conditions, by providing a proper plus clearance to the first and second ball bearings 5a and 6a, d _m n is 70 to 2,000,000, and further d _m n is 80 to 300. It is possible to continuously perform ultra-high speed rotation, as was said.

Table 1

[Table 1]

Among the above conditions, the rotation speed (d _m
n) and the ratio of the radial load F _r to the axial load F _a (F _r / F _a ), these elements (d _m n, F _r /
When the effect of F _a ) on the fatigue life of the ball bearing device was calculated, the results shown in FIG. 6 were obtained. As a precondition for this calculation, the contact angle of the first ball bearing 5a on the load side was 30 degrees, and the contact angle of the second ball bearing 6a on the anti-load side was 15 degrees. In FIG. 6, the X-axis represents the ratio (F _r / F _a ) of the radial load F _r and the axial load F _a to the Y-axis.
The axis represents the rotation speed (d _m n), the Z axis represents the ratio of the life (L / L ₀ ) to the life L ₀ of the conventional structure shown in FIGS.
Each represents. As is evident from FIG. 6, according to the ball bearing apparatus of the present invention, the rotational speed (d _m n) is from 70 to 2,000,000, a load ratio (F _r / F _a) is 2 or less (F _r
A remarkable effect is exhibited in the range of ≤2F _a ).

As is clear from the above calculation results, in the ball bearing device in which the contact angle α ₂ of the second ball bearing 6a on the anti-load side is small, there is a slip between the ball 12 and the outer ring raceway 14 and the inner ring raceway 15. Unless it occurs, the life of the bearing is significantly longer than that of the conventional ball bearing device. However, the life shown in the above table and FIG. 6 is the same as that of the ball 12 which constitutes the second ball bearing 6a on the anti-load side.
The numerical values in the case where no slippage occurs between the rolling surface and the outer ring raceway 14 and the inner ring raceway 15 are shown. As long as this slippage does not occur, the contact angle α ₂ of the second ball bearing 6a on the counter-load side is made small, and a positive clearance is given to the first and second ball bearings 5a, 6a. It is possible to significantly extend the life as compared with the conventional device shown in FIGS. However, when such a small contact angle α ₂ and a positive clearance are combined, a slip often occurs between the rolling surface of the ball 12 and the outer ring raceway 14 and the inner ring raceway 15, and rather the ball bearing device is rather slipped. As described above, the life may be significantly shortened.

Therefore, in the case of the ball bearing device of the present invention,
By limiting the radius of curvature r _e2 of the outer ring raceway 14 of the second ball bearing 6a and the radius of curvature r _i2 of the inner ring raceway 15 in relation to the outer diameter d of the ball 12, it is possible to shorten the life. Prevents slippage. That is, in the case of the ball bearing device of the present invention, as shown in FIG. 7, the radius of curvature r of the cross-sectional shape of the outer ring raceway 14 of the outer ring 13 forming the second ball bearing 6a.
The relationship between _e2 , the radius of curvature r _i2 of the cross-sectional shape of the inner ring raceway 15 of the inner ring 10 that constitutes the second ball bearing 6a, and the outer diameter d of the ball 12 that constitutes this second ball bearing 6a, 0.53d ≦ r _e2 ≦ 0.56d and 0.505d ≦ r _i2 ≦ 0.52d.

By controlling the radius of curvature r _e2 , r _i2 of the cross-sectional shape of each of the raceways 14, 15 within the above range, it is possible to prevent slippage between these raceways 14, 15 and the rolling surface of the ball 12. This will be described with reference to FIG. The chain line in FIG. 7 shows the cross-sectional shape of the outer ring raceway 14a of a conventionally known angular type ball bearing. In the case of this general ball bearing, the radius of curvature r _e2 of the cross-sectional shape of the outer ring raceway 14a
′ Is the outer ring raceway 1 according to the present invention shown by the solid line in FIG.
4 is smaller than the radius of curvature r _e2 of the sectional shape (r _e2 ′ <r
_e2 ).

On the other hand, the rolling surface of the ball 12 and each of the orbits 14 described above.
When a positive gap is provided between a and 15, the balls 12 are displaced diametrically outward (upward in FIG. 7) due to centrifugal force. Then, based on this displacement, the contact point P between the rolling surface of the ball 12 and the outer ring raceway 14a becomes
It moves to the outside of a (the part where the inside diameter becomes large). In the case of the conventional shape shown by the chain line, the contact angle between the ball 2 and the outer raceway 14a becomes 0 degree when the contact point P is displaced from the normal position by l. In such a state where the contact angle is 0 degree, the outer ring raceway 14a has almost no force for pressing the balls 12 toward the inner ring raceway 15. Therefore, the contact pressure between the rolling surface of the ball 12 and the inner ring raceway 15 is almost eliminated, and the ball 12 is hardly pressed while being pressed against the outer ring raceway 14a by the centrifugal force. As a result, abnormal sliding occurs between the rolling surface of the ball 12 and the inner ring raceway 15, and the rolling surface and the inner ring raceway 1
5 produces significant wear.

On the other hand, in the case of the present invention, the radius of curvature r _e2 of the sectional shape of the outer ring raceway 14 on the inner peripheral surface of the outer ring 13 which constitutes the second ball bearing 6a is 0.53d to 0.56d. The radius of curvature r _i2 of the cross-sectional shape of the inner ring raceway 15 (= 0.505d
.About.0.52d) (r _e2 > r _i2 ).
Therefore, the above-mentioned remarkable wear due to slippage does not occur.
That is, as the radius of curvature _re2 is increased, the balls 12
For the contact angle of 0 to become 0 degree, the contact point P must be displaced from the normal position by L (> l). On the other hand, since the radius of curvature r _i2 of the sectional shape of the inner ring raceway 15 is relatively small, the rolling surface of the ball 12 and the inner ring raceway 15 contact each other before the ball 12 is displaced by L. As a result, even when the balls 12 are displaced outward in the diametrical direction by centrifugal force, the contact state between the rolling surface of the balls 12 and the outer ring raceway 14 and the inner ring raceway 15 is ensured, and the contact is made. A sufficient surface pressure acts on the part. Therefore, the balls 12 roll with the operation of the ball bearing device. As a result, it is possible to prevent the occurrence of slippage between the rolling surface of the ball 12 and the inner ring raceway 15, and it is possible to reliably prevent the occurrence of remarkable wear as described above.

As is apparent from the above description, the second ball bearing 6a which is operated in an unloaded state with a positive clearance provided.
Further, in order to prevent the occurrence of slippage that may lead to remarkable wear, the radius of curvature r _i2 of the cross section of the inner ring raceway 15 may be made small and the radius of curvature r _e2 of the cross section of the outer ring raceway 14 may be made large. However, the radius of curvature of the raceway surface must be secured to some extent (0.505d or more) in relation to the outer diameter d of the ball 12 in order to smoothly roll the ball 12. Therefore, the radius of curvature r _i2 of the cross section of the inner ring raceway 15 is also 0.50.
It must be 5d or more. On the other hand, if the radius of curvature of the raceway surface is made too large, the contact area between the raceway surface and the rolling surface of the ball 12 becomes narrow, and when a load is applied to the second ball bearing 6a or centrifugal force is applied. When the ball 12 is pressed against the outer ring raceway 14 on the basis of this, the surface pressure applied to the contact portion becomes too large, which causes the life of the second ball bearing 6a to be shortened. Therefore, by increasing the curvature radius r _e2 of the cross section of the outer ring raceway 14 beyond 0.56d it is not preferable.

From the above description, the radius of curvature r _i2 of the sectional shape of the inner ring raceway 15 is 0.505 d or more (r _i2 ≧ 0.505).
d) and the radius of curvature r _e2 of the cross section of the outer ring raceway 14 is 0.5.
6d below (r _e2 ≦ 0.56d) and it can be seen that must be. Further, in order to prevent the occurrence of slippage, it is necessary to make the radius of curvature r _i2 of the cross section of the inner ring raceway 15 sufficiently smaller than the radius of curvature r _e2 of the cross section of the outer ring raceway 14 (r _i2 <r _e2 ). Things are also clear from the above description. For these reasons, the respective radii of curvature r _i2 and r _e2 must be restricted within the following range.

0.505d ≦ r _i2 ≦ 0.52d 0.53d ≦ r _e2 ≦ 0.56d More preferably, 0.505d ≦ r _i2 <0.51d 0.53d <r _e2 ≦ 0.56d. The most preferable value is r _i2 = 0.505d re ₂ = 0.54d. By restricting the respective radii of curvature r _i2 and r _e2 in such ranges, the contact surface pressure of the contact portion between the rolling surface of the ball 12 and the inner ring raceway 15 and the outer ring raceway 14 is suppressed to a low level as described above. It is possible to prevent abnormal wear due to such slippage.

Furthermore, by controlling the radius of curvature of the cross section of the inner ring raceway 15 and the outer ring raceway 14 of the first ball bearing 5a which rotates while receiving a load during operation, a positive clearance is provided to operate in an unloaded state. The possibility that slippage will occur in the second ball bearing 6a can be further reduced. The reason for this will be described with reference to FIG. The slip causes the balls 12 forming the second ball bearing 6a to be displaced toward the large diameter side of the outer ring raceway 14 based on the centrifugal force, and the contact pressure between the rolling surface of the ball 12 and the inner ring raceway 15 is increased. It occurs at zero or near zero. Therefore, in order to prevent the occurrence of the slip, it is necessary to displace the inner ring raceway 15 on the outer peripheral surface of the inner ring 10 constituting the second ball bearing 6a in the direction away from the balls 12 (leftward in FIG. 5). Preventing is effective. On the other hand, in the inner ring 10, when the axial load applied to the first ball bearing 5a is large, the rolling surface of the ball 11 forming the first ball bearing 5a, the outer ring raceway 14 and the inner ring raceway 15 are separated from each other. Based on the elastic deformation of the abutting portion, it is displaced leftward in FIG. As is clear from this, suppressing the elastic deformation of the first ball bearing 5a operated under load is effective in preventing slippage of the second ball bearing 6a operated under no load.

In order to suppress the elastic deformation to a minimum, the radius of curvature r _e1 of the outer ring raceway 14 of the first ball bearing 5a and the radius of curvature r _{i1 of} the inner ring raceway 15 are
It should be close to 1/2 of the outer diameter d '(d' = d in the case of this embodiment) of the ball 11 constituting the first ball bearing 5a (however, 0.505d 'or more for the above-mentioned reason). Is effective. However, if both of the above-mentioned radii of curvature r _e1 and r _i1 are brought close to 1/2 of the above-mentioned outer diameter d ′, the central axis of the outer ring 13 and the inner ring 1
When the center axis of 0 is inclined, an unreasonable force is applied to the contact portion between the rolling surface of the ball 11 and the outer ring raceway 14 and the inner ring raceway 15, and the first ball bearing 5a is easily damaged. As described above, in the case of the screw compressor in which the ball bearing device of the present invention is incorporated, the rotary shaft 2 fitted with the inner ring 10 is externally fitted.
(Fig. 3) tends to tilt, so in order to avoid such a situation,
It is necessary to increase the curvature radii r _e1 and r _i1 to some extent.

Therefore, in the case of the present embodiment, while preventing the first ball bearing 5a from being damaged due to the inclination of the rotary shaft 2, the inner ring 10 constituting the first ball bearing 5a has the structure shown in FIG. In order to prevent displacement to the left, the above radius of curvature r _e1 ,
r _i1 is regulated so as to satisfy the following conditions (1) to (3). r _i1 ≦ r _e1 −−− (1) r _i1 ≦ 0.52d ′ −−− (2) r _i1 + r _e1 ≧ 1.03d ′ −−− (3)

Among the above conditions (1) to (3), (1)
The condition (2) is that the elastic deformation of the contact portion between the rolling surface of the ball 12 and the inner ring raceway 15 is made small (the contact area of the contact portion is made wide), and the inner ring 10 is moved to the left in FIG. It is necessary to prevent the displacement. The condition (3) is necessary to prevent the first ball bearing 5a from being unduly stressed when the rotary shaft 2 is tilted. The outer ring raceway 14 of the second ball bearing 6a which has a small contact angle α _{2 and} is operated under no load
And not only the radius of curvature r _e2 , r _i2 of the cross-sectional shape of the inner ring raceway 15 is restricted as described above, but also the curvature of the cross-sectional shape of the first ball bearing 5a that has a large contact angle α _{1 and} is operated under load. By limiting the radii r _e1 and r _i1 as described above, the life of the ball bearing device formed by combining these first and second ball bearings 5a and 6a can be made significantly longer than that of the conventional ball bearing device. Can be extended. Further, compared to the ball bearing device described in each of the above publications, the effect of extending the life is ensured.

FIG. 8 shows the result of calculation performed to confirm the effect of the present invention. The preconditions of the calculation are the same as the conditions when the above Table 1 was prepared except that the axial load was 100 kgf. Further, the contact angle of the first ball bearing 5a on the load side was 30 degrees, and the contact angle of the second ball bearing 6a on the anti-load side was 15 degrees. In FIG. 8, the horizontal axis represents the axial clearance (effective axial clearance) during operation of the ball bearing device formed by combining the first and second ball bearings 5a and 6a, and the vertical axis represents the calculated life. ing. Further, the solid line A is the calculated service life of the embodiment of the present invention, and the broken line B is an example having a curvature outside the limit of the present invention.
The calculated lifespans of the comparative examples, which require attention when implemented, are shown.

The first ball bearing 5a of the embodiment of the present invention represented by the solid line A and the comparative example represented by the broken line B are shown.
And the outer ring raceway 14 and the inner ring raceway 15 of the second ball bearing 6a.
The radius of curvature of was regulated as follows. The outer diameters d and d'of the balls 11 and 12 are the same (d = d '). Example First ball bearing 5a r _i1 = 0.505d 'r _e1 = 0.525d' Second ball bearing 6a r _i2 = 0.505d r _e2 = 0.540d Comparative example First ball bearing 5a r _i1 = 0.515d 'r _e1 = 0.515d' Second ball bearing 6a r _i2 = 0.515d r _e2 = 0.515d

As is clear from the description of FIG. 8, the ball bearing device of the present invention can significantly extend the life as compared with the conventional structure. Moreover, since a long life can be realized with a sufficient effective axial gap provided, it is possible to reliably support a rotating shaft that rotates at an ultra-high speed while often changing the tilt angle, like a rotating shaft for a screw compressor. That is, according to the embodiment of the present invention, sufficient durability (calculated life) can be obtained even if the effective axial gap is relatively large, for example, about 5 to 25 μm.

On the other hand, in the case of the comparative example, not only the durability originally obtained is inferior to that of the example, but also when the effective axial gap exceeds 3 to 6 μm, the above-mentioned slippage occurs. Durability will drop sharply. Therefore, at the time of use, it is necessary to take care so that the effective axial gap falls within the range of 0 to 6 μm. As described above, when the radius of curvature deviates from the present invention, the range of the value of the effective axial gap in which the durability sharply decreases becomes smaller, and it becomes practically practically useless.

Incidentally, the first ball bearing 5 which constitutes the ball bearing device.
relationship between a calculation lifetime L _5a and calculated life L _6a of the second ball bearing 6a is that the 3L _5a ≦ L _6a is, these two ball bearings 5
Calculated life L _{T of} ball bearing device consisting of a and 6a
It is preferable from the viewpoint of securing. The reason for this is as follows. The calculated life of a ball bearing device consisting of a combination of multiple ball bearings does not match the calculated life of the ball bearing with the shortest calculated life (short-life bearing), but becomes shorter than this.
Being affected by the calculated life of ball bearings (long-life bearings) with a long calculated life, and further, the longer the calculated life of long-life bearings,
It is conventionally known that the calculated life L _T of a ball bearing device approaches the calculated life of a short-life bearing.

On the other hand, while it is difficult to extend the calculated life L _5a of the first ball bearing 5a that receives an axial load during use, the second ball bearing 6 that does not receive such an axial load is difficult.
It is easy to lengthen the calculation life L _6a of a. Therefore,
In the case of the ball bearing device of the present invention, the calculated life L _5a of the first ball bearing 5a is considered to be the calculated life of the short-life bearing, and the calculated life L _6a of the second ball bearing 6a is the calculated life of the long-life bearing. Can be thought of. Therefore, these first and second ball bearings 5a,
6a the calculated life L _5a, the ratio of L _6a and (L _6a / L _5a), the calculated life of the calculated life L _T and the first ball bearing 5a of both ball bearings 5a, ball bearing apparatus comprising a combination of 6a When showing the relationship between the specific (L _T / L _5a) of the L _5a, it becomes as in FIG.
This FIG. 9 is drawn based on a conventionally known formula for calculating the life of a combined bearing. As is apparent from FIG. 9, if the calculated life L _6a of the second ball bearing 6a is set to be three times or more the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the ball bearing device will be the first. Can be 80% or more of the calculated life L _5a of the ball bearing 5a.

Of course, if the calculated life L _6a of the second ball bearing 6a is made significantly longer than the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the ball bearing device will be Ball bearing 5
It can be brought closer to the calculated life L _5a of a. However,
Even if the calculated lifespan L _6a of the second ball bearing 6a is greatly extended over three times the calculated lifespan L _5a of the first ball bearing 5a, the cost required to extend the calculated lifespan L _6a is The effect of extending the calculated life L _T of the ball bearing device obtained thereby is small. Therefore, considering the total cost, it is appropriate that the calculated life L _6a of the second ball bearing 6a exceeds three times the calculated life L _5a of the first ball bearing 5a.

Next, FIG. 10 shows a second embodiment of the present invention. In the case of this embodiment, the first and second ball bearings 5a, 6
The balls 11 and 12 of a are held rotatably by the cages 16 and 16. The cages 16, 16 are so-called machined cages, and are pockets 18, 1 for rollingly holding the balls 11, 12 in a cylindrical main portion 17, 17.
8 is formed. Such cages 16 and 16 are the outer ring 1 which comprises the 1st and 2nd ball bearings 5a and 6a, respectively.
The outer ring guides 3 and 13 are mounted inside. Particularly, in the illustrated embodiment, the outer peripheral surfaces of both ends of the retainers 16 and 16 in the axial direction (the left-right direction in FIG. 10) and the outer rings 13 and 13 are shown.
Both shoulder portions of the inner peripheral surface, that is, the portions deviated from the outer ring raceways 14, 14 are opposed to each other with a minute gap therebetween to form a so-called outer ring shoulder guide.

The reason why the guide state of each of the cages 16 and 16 is the outer ring guide is as follows. That is,
The general use condition as a ball bearing device, that is, d _m n is 7
If the preload is applied to the first and second ball bearings 5a and 6a and the vibration is less generated during operation, the guide conditions for the cages 16 and 16 need not be particularly restricted. Absent. On the other hand, in the case of the ball bearing device incorporated in the screw compressor, which is the object of the present invention, d _m n is 70 or less.
It operates in the ultra-high speed range of over 10,000 with a positive axial clearance. Moreover, the vibration increases due to the fluctuation of the axial load peculiar to the screw compressor. In order to suppress the wear of the guide surfaces of the cages 16 and 16 in such a usage state, it is necessary to use outer ring guides that can suppress the inclination of the cages 16 and 16.

On the other hand, when these cages 16, 16 are used as rolling element guides or inner ring guides, the rigidity is lower than that of the inner ring 10 and the outer ring 13 based on centrifugal force during operation. ) As the diameters of the cages 16, 16 expand elastically, the spacing between the guide surfaces increases, and it becomes impossible to provide sufficient guidance. When vibrations are applied to the cages 16 and 16 during operation in this state, the surfaces of the cages 16 and 16 and the mating surface come into contact with each other in a small area, which causes inconvenience such as abnormal wear. When the outer ring guide is used, the cages 16, 1 are accompanied by centrifugal force.
Even if the diameter of 6 tends to widen, the guide surface (the cage 16,
The distance between the outer peripheral surface of 16 and the inner peripheral surfaces of the outer rings 13 and 13 does not increase. Further, since the lubricating oil exists between the guide surfaces, the guide surfaces do not rub each other directly. Therefore, by using the cages 16 and 16 as outer ring guides, abnormal wear of the guide surfaces can be effectively prevented. Incidentally, the cages 16, 1
If a retainer made of copper alloy is used as 6, retainer 1
The strength of the retainers 16 and 16 can be effectively increased by increasing the strength of the retainers 16 and 16.

In addition, a plus clearance is provided, a contact angle of the second ball bearing 6a is made smaller than a contact angle of the first ball bearing 5a, and outer ring raceways 14, 14 of the ball bearings 5a, 6a.
Also, the point that the radius of curvature of the cross-sectional shape of the inner ring raceways 15, 15 is restricted is the same as in the first embodiment described above. If the d _m n in use is not so high and therefore the centrifugal force applied to the cages 16 and 16 is limited, as in the third embodiment shown in FIG. May be used as one-shoulder guide for the outer ring.

Next, FIG. 12 shows, as a fourth embodiment of the present invention, a more specific structure of the supporting portion of the rotary shaft of the screw compressor. Between the outer peripheral surface of the rotary shaft 2 and the inner peripheral surface of the housing 3, a non-contact type seal such as a labyrinth seal or a mechanical seal is arranged in this order from the side of the rotor 1 housed in the housing 3. A seal device 19 such as a contact type seal, a roller bearing 4 for bearing a radial load, a first ball bearing 5a and a second ball bearing 6a constituting the ball bearing device of the present invention are connected in series with each other. It is installed in. A spacer 7 is provided between the inner ring 20 of the roller bearing 4 and the first ball bearing 5a, and a spacer 7 is provided between the outer ring 21 of the roller bearing 4 and the outer ring 13 of the first ball bearing 5a. The nozzle rings 22 are respectively sandwiched.

The inner peripheral portions on both side surfaces of the nozzle ring 22 are inclined toward the inner peripheral edge so as to approach each other.
It is a conical concave inclined surface. The nozzle holes 23, 23 are opened at one or a plurality of positions on the both inclined surfaces in the circumferential direction. Each nozzle hole 23, 23 is formed in a direction perpendicular to the inclined surface. The inclined surface is provided in order to facilitate the work of forming the nozzle holes 23, 23 in the inclined direction. Therefore, each of these nozzle holes 23,
The opening 23 is directed inwardly in the diametrical direction. Then, in each of these nozzle holes 23, 23, an oil supply passage 24 formed in the housing 3 and the nozzle ring 22 is provided.
The lubricating oil can be fed in through the.

On the other hand, the above-mentioned first and second ball bearings 5a, 6a
Of the cages 16a, 16a incorporated in the
The inner peripheral surface of both end portions is a conical concave inclined surface whose inner diameter increases toward the edge portion. Therefore,
The width of the space existing between the outer peripheral surfaces of the inner rings 10 and 10 forming the ball bearings 5a and 6a and the inner peripheral surfaces of the retainers 16a and 16a in the diametrical direction (vertical direction in FIG. 12). The dimension increases toward the open end. As a result, the efficiency of feeding the lubricating oil ejected from the nozzle hole 23 into the first and second ball bearings 5a and 6a is improved. Therefore, even when the ball bearing device composed of these ball bearings 5a and 6a is operated at ultra-high speed, these ball bearings 5
It is possible to take in a sufficient amount of lubricating oil in a and 6a.

The reason why the direction in which the nozzle hole 23 is opened and the shape of the inner peripheral surface of each of the cages 16a, 16a are set as described above is as follows. In other words, the ball bearing device of the present invention can greatly extend the life as compared with the conventional device described above. Therefore, if the required life is greatly exceeded when it is made in the same size as the conventional one, the size series of the first and second ball bearings 5a and 6a should be reduced so that the inner diameter does not change. Can be reduced. This makes it possible to reduce the size and cost of the ball bearing device, but a ball bearing with a small size series has a small internal space. For this reason, jet refueling generally performed in screw compressors (breaking the wall of air generated inside the bearing based on high speed rotation,
It becomes difficult to carry out the lubrication method by sending lubricating oil into the inside of the bearing. Therefore, by adopting the above-described opening direction of the nozzle holes 23, 23 and the shape of the inner peripheral surface of the retainers 16a, 16a, the jet oil supply can be performed.

The inclination angle of the nozzle hole 23 with respect to the rotary shaft 2 is optimally about 15 degrees, but can be set within the range of 10 to 20 degrees. Further, the nozzle ring 2 having the nozzle hole 23 and also having a function as an outer ring spacer.
2 has a hardness of HRc56 to 66 (more preferably HRc60
Preferably, it is made of a steel material in the range of -62). This is because both sides of the nozzle ring 22, which come into contact with each other, and the outer races 21, 1 of the roller bearing 4 and the first ball bearing 5a are brought into contact with each other due to the peculiar vibration generated when the screw compressor is operated.
This is to prevent fretting wear from occurring with the end face of No. 3.

In addition, a plus clearance is provided, the contact angle of the second ball bearing 6a is made smaller than the contact angle of the first ball bearing 5a, and the outer ring raceways 14, 14 of each ball bearing 5a, 6a.
The point that the radius of curvature of the cross-sectional shape of the inner ring raceways 15, 15 is restricted is the same as in the above-described first embodiment and the above-described second to third embodiments.

FIG. 13 shows a fifth embodiment of the present invention. In the case of the present embodiment, the second ball bearing 6 on the non-load side
The diameter of the ball 12a forming a is smaller than that of the ball 11 forming the load-side first ball bearing 5a. The smaller the diameter of the ball 12a, the smaller the internal axial load generated inside the ball bearing device due to the centrifugal force applied to the ball 12a during use, and the durability of the ball bearing device can be further improved. Although the load capacity is reduced by using the small diameter ball 12a, the load applied to the second ball bearing 6a is small, and therefore, even when the small diameter ball 12a is used, practically sufficient durability can be secured. In addition, a point where a positive gap is provided, a point where the contact angle of the second ball bearing 6a is made smaller than the contact angle of the first ball bearing 5a, the first and second ball bearings 5a, 6
The point that the radii of curvature of the cross-sectional shapes of the outer ring raceways 14 and 14 and the inner ring raceways 15 and 15 of a are regulated is the same as in each of the above-described embodiments.

Although not shown, the number of balls forming the second ball bearing on the anti-load side is smaller than the number of balls forming the first ball bearing on the load side. The internal axial load can also be reduced. Although the load capacity is reduced by reducing the number of balls, since the load applied to the second ball bearing is small, practically sufficient durability can be secured even when the number of balls is reduced. In this way, when the number m of the balls forming the second ball bearing is set to be smaller than the number n of the balls forming the first ball bearing (m <n), the small number m is set to 7
It is preferable to regulate to 0 to 80% (m = (0.7 to 0.8) n). When incorporating more than 80% of balls, the effect of reducing the internal axial load becomes insufficient. On the other hand, when it is less than 70%, the distance between the adjacent balls becomes too wide, which hinders smooth rotation.
In this way, the technique of reducing the number of balls forming the second ball bearing on the anti-load side to less than the number of balls forming the first ball bearing on the load side is implemented in combination with the present invention. Other than that,
It is obvious that the present invention can be carried out independently of the present invention.

Similarly, the internal axial load can be reduced by making only the balls constituting the second ball bearing on the counter-load side, which are lighter in weight than the bearing steel. By using a ceramic ball, the surface pressure applied to the raceway surface that contacts the rolling surface of the ball increases as described above, but the load applied to the second ball bearing is small, so an excessive surface pressure is applied. It does not work and can ensure sufficient durability for practical use. Of course, in this way, the diameter of the balls that make up the ball bearing on the anti-load side can be reduced, the number of balls can be reduced, and the technology of using ceramic balls can be used alone or in combination. You can also In this way, as a result of making the internal axial load smaller, the radial load F _r is ⅕ or less of the axial load F _a (5F _r ≦ F) at a higher speed such as d _m n of 80 to 300. _It can be used under more severe conditions such as _a ).

In each of the illustrated embodiments, a pair of angular type ball bearings is combined with DF, but a plurality of first ball bearings that support the axial load F _a during operation are shown. It is also possible to provide the plurality of first ball bearings in parallel combination with each other and one second ball bearing that does not support the axial load Fa to form _a DF.

[0066]

Since the present invention is constructed and operates as described above, it is possible to improve the durability of the ball bearing device by reducing the internal axial load while ensuring a sufficient load capacity against the external axial load. Can be achieved. Moreover, the tolerance for the inclination of the rotary shaft is large, and abnormal wear due to slipping can be reliably prevented, so that superior durability can be reliably obtained.

[Brief description of drawings]

FIG. 1 is a sectional view showing a first example of a conventional device.

FIG. 2 is a sectional view of the same main part.

FIG. 3 is a sectional view showing the second example.

FIG. 4 is a sectional view of the same main part.

FIG. 5 is a sectional view similar to FIG. 4, showing a state where slippage occurs in the ball bearing on the anti-load side.

FIG. 6 is a diagram showing the influence of the rotational speed and the ratio of the axial load to the radial load on the life of the ball bearing device.

FIG. 7 is a cross-sectional view showing a cross section of a second ball bearing that constitutes the ball bearing device of the present invention with exaggerated radius of curvature of the raceway.

FIG. 8 is a diagram showing the influence of the radius of curvature of the raceway and the axial clearance on the calculated life of the ball bearing device.

FIG. 9: Ratio of the calculated life of the first ball bearing and the calculated life of the second ball bearing, and the ratio of the calculated life of the ball bearing device formed by combining these and the calculated life of the first ball bearing. FIG.

FIG. 10 is a cross-sectional view of a main part showing a second embodiment of the present invention.

FIG. 11 is a cross-sectional view of a main part showing the third embodiment.

FIG. 12 is a cross-sectional view of an essential part showing the fourth embodiment.

FIG. 13 is a cross-sectional view of a main part showing the fifth embodiment.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 rotor 2 rotating shaft 3 housing 4 roller bearing 5, 5a first ball bearing 6, 6a second ball bearing 7, 8 spacer 9 retaining ring 10 inner ring 11 ball 12, 12a ball 13 outer ring 14, 14a outer ring raceway 15 Inner ring track 16, 16a Cage 17 Main part 18 Pocket 19 Sealing device 20 Inner ring 21 Outer ring 22 Nozzle ring 23 Nozzle hole 24 Oil supply passage

Claims (2)

[Claims] 1. A first and a second ball bearing, which are provided between an outer peripheral surface of a shaft and an inner peripheral surface of a housing, are angular type and have different contact angle directions from each other. In a ball bearing device for supporting an axial load applied from the outside in a substantially constant direction between the shaft and the housing by a first ball bearing having a contact angle larger than that of the second ball bearing, The first and second ball bearings are arranged in a front combination and each has a positive gap. The radius of curvature of the cross-sectional shape of the outer ring raceway of the second ball bearing is r
_{where e2} and the outer diameter of the ball forming the second ball bearing are d, 0.53d ≦ r _e2 ≦ 0.56d is satisfied, and the curvature of the cross-sectional shape of the inner ring raceway of the second ball bearing is satisfied. Radius r
_i2 and then, the outer diameter of the ball configuring the second ball bearing in the case of the d, ball bearing apparatus characterized in that satisfy 0.505d ≦ r _i2 ≦ 0.52d. 2. When the radius of curvature of the cross-sectional shape of the outer ring raceway of the first ball bearing is r _e1 and the outer diameter of the balls forming the first ball bearing is d ′, 0.505 d ′ ≦ r _e1 ≦ 0.53d ′ is satisfied, and the radius of curvature of the cross-sectional shape of the inner ring raceway of the first ball bearing is r
_i1 and the outer diameter of the ball forming the first ball bearing is d ′, 0.505d ′ ≦ r _i1 ≦ 0.52d ′ is satisfied, and further, r _e1 + r _i1 ≧ 1.03 d ′ The ball bearing device according to claim 1, which satisfies: