JP2001280347A - Ball bearing - Google Patents

Abstract

PROBLEM TO BE SOLVED: To increase a supply amount of the grease to a raceway surface, and elongate the lubricating life of a capped grease-included bearing. SOLUTION: In this ball bearing, an inner face of a pocket 2 of a cage 1 is a spherical face. The inner face of the pocket 2 of the cage 1 has a radial groove 4 penetrating from an inside cylindrical surface to an outside cylindrical surface of the cage. This groove 4 has the circular arc-shaped cross section.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a spindle motor for HDDs, various fan motors, general-purpose motors, and the like. Related to ball bearings.

[0002]

2. Description of the Related Art A sealed ball bearing is generally used as a ball bearing for supporting a rotating shaft of a motor, and a rigidity of the bearing is secured by applying a preload in an axial direction. In addition, high rotational accuracy of the rotating body is maintained.
In a sealed ball bearing, grease is sealed inside the bearing, and the life of the bearing is determined by the life of the grease due to repeated shearing in a contact surface that receives an axial load on the ball and the raceway surface. Therefore, the performance of grease has been conventionally improved.

In a mechanical device using a sealed ball bearing, extending the life is an eternal problem, and the life of the bearing largely depends on the life of the grease as described above. In a ball bearing in which an axial load is applied by a motor or the like, generally, the raceway surface and the area of about 1 mm ^{2 of the} ball are elastically deformed and the load is supported by a grease film of about 1 μm. The grease moves between the inner and outer raceway surfaces as the ball rotates. At this time, the grease is scraped off by the edge of the cage in which the balls are arranged in the circumferential direction, so that the amount of grease supplied to the contact area between the balls and the raceway decreases with the rotation of the bearings. Bearings are used. When the amount of grease supplied decreases, the grease film supporting the load becomes thinner.If the grease film is thinner than the roughness of the raceway surface, the ball comes into contact with the raceway surface, and the surface of the ball and the raceway surface wear out. As a result, the vibration of the bearing increases. Therefore, in a hard disk drive, an air conditioner fan motor, and the like that require quietness, the bearing life is early. Further, the heat generated by the contact between the ball and the raceway accelerates the oxidative deterioration of the grease. In addition, the wear powder acts as a catalyst for accelerating the oxidative deterioration of the grease, thereby shortening the lubrication life of the bearing.

In order to prevent the grease from being scraped off by the edge of the outer diameter surface of the retainer, a method has been proposed in which a concave portion is provided at the edge of the pocket opening (for example, JP-A-11-166540). ). In this proposal, a chamfered concave portion is provided at a part of the pocket opening edge on the inner peripheral surface and the outer peripheral surface of the ring-shaped retainer to improve the flowability of grease into the retainer pocket surface.
As the concave portion, instead of a chamfered shape, a concave groove extending from the inner peripheral surface to the outer peripheral surface of the retainer, and an intermediate portion of the concave groove portion being substantially flush with the pocket inner surface has also been proposed. ing. However, on the inner surface of the pocket, scraping of grease occurs at the contact portion between the ball and the inner surface of the pocket,
There has been a problem that the amount of grease passing through the pocket inner surface cannot be sufficiently increased. In addition, grease on the ball surface is present on both sides of the Hertzian contact width under load, and there is a problem that the grease is not scraped off and cannot contribute to lubrication.

[0005] In these motors, low torque is required for bearings used for the purpose of reducing power consumption as a mechanical device.

[0006] The ball which is rotating on the inner surface of the pocket of the cage generates sliding friction on the contact surface thereof, which is a factor in generating torque of the bearing itself.

[0007] It is an object of the present invention to increase the amount of grease supplied to a lubricating surface and extend the lubrication life of a bearing.
Another object of the present invention is to reduce the sliding friction generated on the inner surface of the pocket of the retainer and achieve a lower torque of the bearing.

[0008]

A ball bearing according to the present invention holds a plurality of balls interposed between an inner ring and an outer ring in a pocket provided in a ring-shaped retainer, and forms an inner surface of the pocket in a spherical shape. In the ball bearing described above, a radial groove penetrating from the inner diameter surface to the outer diameter surface of the cage is provided on the inner surface of the pocket of the cage, and the cross-sectional shape of the groove is arc-shaped. The groove has a depth equal to or greater than a predetermined value over the entire length of the groove, that is, there is no portion where the groove depth becomes zero. Since the groove penetrating in the radial direction is provided on the inner surface of the retainer, a space is formed at a location where the lubricant such as grease attached to the ball passes through the inner surface of the pocket. This prevents the opening edge of the pocket or the inner surface of the pocket from being scraped off, and allows a sufficient amount of lubricant to move to the other raceway surface. Since the radial groove has an arc-shaped cross-section and the center of the groove width is deep and the side edges are shallow, lubricant is effectively supplied to the ball from the groove in the pocket. You. Therefore, the amount of grease supplied to the raceway surface can be increased, and the lubrication life of the sealed grease-enclosed bearing can be extended.

In the present invention, the axial position of the groove of the retainer is preferably a position including the center of the ball within the groove width. In general, in a ball bearing, the raceway surfaces of the inner and outer rings have an arc-shaped cross section slightly larger than the ball diameter, and the ball contacts a part of the width of the raceway surface. In addition, a contact angle occurs due to an axial load acting on the bearing, and the contact occurs at a position shifted in the axial direction from the center of the pitch circle of the ball arrangement. During the rotation of the bearing, a portion with a large amount of lubricant is generated in the width direction portions on both sides of the ball contact point on the raceway surface. Therefore, if the position of the radial groove of the retainer is set at the axial position including the center of the ball, the grease adhering to the ball can be effectively moved to the location requiring lubrication.
It is even more preferable that the position of the radial groove coincides with the axial position of the center of the ball. As described above, when the center of the ball is included in the groove width, the point of contact of the inner and outer raceways with the ball on the raceway surface is set to the diameter when the axial preload is applied between the inner race and the outer race. It is more preferable to set the position and the groove width of the groove so as to be located within the groove width of the groove in the direction. With this setting, the grease adhering to the ball can be more effectively moved to the location requiring lubrication.

In the present invention, the groove width dimension of the groove is
More preferably, the width is equal to or larger than the width of the hertz contact surface. That is, each Hertzian contact ellipse major axis radius a _A between the balls and the inner and outer rings of the raceway surface calculated from the applied load and the bearing internal specifications, the larger was the a of a _B, the contact angle theta,
When the outer diameter of the retainer is Dco, the inner diameter of the retainer is Dci, and the pitch circle diameter of the ball is Dp, the axial center of the groove is the axial center of the pocket surface of the retainer, and the axial direction of the groove is The width 2L is calculated by the following formula

[0011]

(Equation 2)

It is preferable to use the larger one of 2X _A and 2X _B defined in the above. The above formula is applied when the ball bearing is a deep groove ball bearing. As described above, when the groove width is wider than the Hertz contact width, grease protruding on both sides of the contact portion can contribute to lubrication without being scraped off on the inner and outer diameter surfaces of the retainer. In addition, since the groove cross section is arc-shaped, the center of the groove width is deep, and both side edges are shallow, so that the grease can be moved to the ball center in the pocket, and the raceway ring hertz contact portion The amount of grease at the center of the inlet increases, and the lubrication life of the sealed grease-filled ball bearing can be improved. The load is, for example, a rated load. The maximum value of the axial width dimension of the groove is, for example, the maximum width that the ball can be held by the remaining inner diameter surface of the retainer even when the groove is formed.

In each of the above configurations of the present invention, in order to increase the amount of grease on the inlet side of the contact portion, the longitudinal cross-sectional shape of the groove bottom surface of the retainer groove is changed to a circle concentric with the spherical pocket inner surface. It may be arc-shaped. The inner surface of the pocket has a spherical shape whose diameter is larger than the diameter of the ball. Therefore, in order to secure the amount of the lubricant that adheres to and moves on the surface of the ball, the vertical cross-sectional shape of the radial groove may be an arc concentric with the inner surface of the pocket.

In each of the above configurations of the present invention, the retainer may be made of resin and may have a crown shape. When a crown-shaped resin cage is used, it can be manufactured at low cost by injection molding or the like, and the number of assembly steps is small, and it is economically excellent.

[0015]

An embodiment of the present invention will be described with reference to the drawings. This ball bearing is composed of a raceway surface 6 of an inner ring 6 and an outer ring
A plurality of balls 5 are interposed between a and 7a, and these balls 5 are held by respective pockets 2 provided in a holder 1. Each of the raceway surfaces 6a and 7a is formed by a circumferential groove having an arc-shaped cross section having a diameter slightly larger than the ball diameter. Seals 8 for sealing the annular space between the inner and outer rings 6, 7 are attached to both sides of the outer ring 7, and this bearing is a sealed ball bearing. Specifically, it is a sealed deep groove ball bearing.
The seal 8 may be a shield plate.

As shown in FIGS. 2 and 3, the retainer 1 is made of a crown and made of resin, and has a pair of claws 3 for each pocket 2 on one width surface of a ring-shaped retainer body 1a. , 3 facing each other. Each pocket 2 is formed by the pair of claws 3 and 3 and a concave portion of the retainer main body 1a, and has an inner surface formed in a spherical shape. This cage 1 is formed by injection molding or the like.

On the inner surface of the pocket 2 of the retainer 1, there is provided a radial groove 4 penetrating from the inner diameter surface to the outer diameter surface.
The groove 4 has a cross-sectional shape, that is, an arc-shaped cross-sectional shape along the axial direction of the cage 1. The degree of the arc may be semicircular or shallower. That is, any shape can be used as long as the amount of grease at the center on the contact surface entrance side can be increased. Note that the cross-sectional shape of the groove 4 may be triangular or elliptical. The position of the groove 4 of the retainer 1 is an axial position including the center O of the ball 5. For example, the center O of the ball 5 matches the center of the groove width of the groove 4. The groove width dimension of the groove 4 is given by
It is preferable that the width be equal to or larger than the width of the hertz contact surface.

As shown in FIG. 4, the inner surface of the pocket 2
The ball 5 has a spherical shape with a diameter slightly larger than the diameter dw. The vertical cross-sectional shape of the bottom surface of the radial groove 4, that is, the cross-sectional shape along the radial direction is an arc concentric with the pocket inner surface, and has a larger diameter (diameter dg) than the pocket inner surface. The value of the diameter dg is designed according to the required amount of grease movement.

According to the ball bearing having this configuration, since the groove 4 penetrating in the radial direction is provided on the inner surface of the cage 1, a space is formed at a location where the grease adhering to the ball 5 passes through the inner surface of the pocket. Become. Therefore, it is prevented that the opening edge of the pocket 2 or the inner surface of the pocket 2 is scraped off,
A sufficient amount of grease can move to the other track surface. That is, in the conventional retainer, the grease adhering to the ball 5 could not be moved to the other raceway surface due to a so-called wiper action of scraping grease at the edges of the inner diameter surface and the outer diameter surface of the claw 3. According to this embodiment, such a problem is solved. This radial groove 4
Has an arc-shaped cross-section, a deep center of the groove width,
Due to the shallow sides, grease is effectively supplied to the ball from the groove 4 in the pocket 2. for that reason,
The amount of grease supplied to the raceway surfaces 6a, 7a can be increased. In addition, since the radial groove 4 has a vertical cross-sectional shape that is concentric with the inner surface of the pocket, the groove 4 is designed to have an appropriate depth, so that the groove 4 adheres to the surface of the ball 5 and moves. The required amount of grease can be secured.

FIG. 1 shows a state in which an axial preload is applied to the bearing from the right side of the outer ring 7 and the left side of the inner ring 6. Because the bearing has a radial internal clearance,
When an axial load is applied, a contact angle θ occurs, and the ball 5 contacts the outer ring 7 at a point A and the inner ring 6 at a point B. Point O is the center of ball 5. Based on the size of the axial load and the size of the radial internal gap, the contact surfaces (contact points A and B) that contact the balls 5 of the inner and outer rings 6 and 7 to support the load
The size and position of the contact surface at the position are determined. The contact surface is a Hertz contact surface. During the rotation of the bearing, a ring-shaped pool of grease is formed on both sides orthogonal to the rotation direction of the contact surface. That is, a portion having a large amount of lubricant is formed. Therefore, in order to effectively move the grease adhering to the ball 5, it is sufficient that the axial center of the radial groove 4 should be coincident with the center point O of the ball 5. In addition, the center point of the radial groove 4 (the center point of the ball 5)
Groove width L from O to pocket open side (claw 3 open side)
_C preferably includes a claw-opening side end C in the width of the contact surface between the outer ring 7 and the ball 5. Similarly, the groove width L _D from the center point O of the radial groove 4 to the pocket non-open side is the pocket non-open side end D in the width of the contact surface between the inner ring 6 and the ball 5.
It is desirable to include

The details of the method for determining the overall groove width dimension 2L of the groove 4 will be described below. By setting the groove width to a region completely including the Hertzian contact width, grease adhering to the surface of the ball 5 and moving can be secured. Groove width 2L, shown below, a case where the axial center of the pocket surface of the cage 1 the axial center of the groove 4, is L = L _C = L _D. As described above, FIG. 1 shows a state in which axial preload (axial load) is applied to the bearing from the right side of the outer ring 7 and the left side of the inner ring 6. Ball 5 is at point A with outer ring 7,
The inner ring 6 contacts each other at a point B. Point O is the center of the ball. The radial internal clearance of the bearing is Cr, the radius of curvature of the raceway surface groove of the inner race 6 is r _B , and the radius of curvature of the raceway surface of the outer race 7 is r.
_A , and assuming that the ball diameter is D _W , the contact angle θ is θ = cos ⁻¹ [(r _A + r _B −D _W / 2) / (r _A +
r _B -D _W )]. The load P applied to the ball is represented by P = Fa / Z according to the axial load Fa and the number of balls Z.
sin θ.

According to the revised lubrication handbook (edited by The Japan Lubrication Society, Yokendo (1987) 16-17), the contact ellipse major axis diameter a _B of the Hertzian contact surface at points B and A,
a _A is, P a load applied to the ball, E the modulus of longitudinal elasticity of the material of the ball, the inner ring or each E _2B modulus of longitudinal elasticity of the outer ring of the material, E _2A, the Poisson's ratio [nu ₁ material ball,
The Poisson's ratio of the material of the inner ring or the outer ring is ν _2B , ν
_2A , the coefficients determined by the principal curvatures of the ball and the inner ring or the ball and the outer ring are A ₁ , A ₂ ,

[0023]

(Equation 3)

Assuming that the principal curvature of the raceway surface of the ball and the inner or outer race and the coefficient obtained as a function of the angle φ formed by the plane including the principal curvature are α _{B and} α _A respectively,

[0025]

(Equation 4)

Is determined. The method for obtaining the alpha _B or alpha _A is due to revised lubricating Handbook (Japan lubricating Gakkai, nourishing Kashikodo (1987) page 16 Figure 2.1.3). In general, the point of contact is that if two curved surfaces make point contact, the contact surface will be elliptical. According to Hertz,
Coefficient α for determining the major axis radius of the contact ellipse as a function of the principal curvature of the two curved surfaces and the angle formed by the plane including the principal curvature
Is obtained.

When the ball 5 rotates, the Hertzian contact surface on the surface of the ball 5 has the contact ellipse major axis diameters a _B , a B of the Hertzian contact surface at the point A and the point B.
_A belt-shaped trajectory is a straight line connecting the contact points A and B with the center passing through the point O with the larger dimension a of A. The areas where the above-mentioned orbits intersect on the cage outer diameter surface and the cage inner diameter surface are the contact angle θ, the width dimension a, the cage outer diameter dimension Dco, the cage inner diameter dimension Dci, and the ball pitch circle diameter. Due to the geometric relationship, including the case where the axial load is in the opposite direction to FIG.

[0028]

(Equation 5)

The larger one of the values 2X _A and 2X _B shown in the above is defined as the width dimension, and the area is determined as the area whose center is the point O. The groove width dimension 2L is set so as to include the above-mentioned area that intersects on the cage outer diameter surface. FIG. 1 shows an example in which an axial load is applied to the bearing. However, the groove width can be designed in the same manner as described above even under a radial load or a combined load of a radial load and an axial load.

In the above embodiment, the retainer 1 is a crown-shaped resin retainer. However, the present invention can be applied to a case where the retainer 1 is a corrugated press retainer.

[0031]

Next, various test results will be described. Regarding the oil film formation test, two types of tests having different test conditions from each other (particularly, significantly different rotation speeds) were performed. [Oil Film Formation Test (1)] (Low-Speed Rotation) In the cage 1 in the embodiment of the present invention, the groove depth D (FIG. 3) is kept constant, and the claw release of the Hertzian contact track on the outer diameter surface of the cage is performed. The ratio of the half width dimension L of the radial groove 4 to the axial dimension X _A (FIG. 1) between the side end and the ball center (L
/ X _A ) was changed, and the performance with the conventional cage having no radial groove 4 was confirmed by an oil film formation test. As a test bearing, a deep groove ball bearing 608 (inner diameter 8 mm × outer diameter 22 mm × width 7 mm) was used. The ball 5 is a steel ball and has a diameter of 4 mm. Grease has a thickener of Li soap and a base oil of synthetic ester oil. Kinematic viscosity of base oil at 40 ° C is 33mm
² / s. Ambient temperature is 80 ° C, axial load is 5
kgf. The test rotation speed was 1050 rpm, and the inner ring was rotated. The cage 1 is an injection-molded nylon crown type cage. A DC circuit for measuring the contact electrical resistance between the stationary wheel and the rotating wheel (inner wheel) during operation was constructed, and the oil film formation rate of the bearing was evaluated from the resistance value. Here, the oil film formation rate is defined as follows from the voltage V between the stationary wheel and the rotating wheel under test with respect to the applied voltage Vo of the bearing. Oil film formation rate = 100 × V / Vo. Note that a mercury slip ring attached to the end of the rotating shaft was used for connection between the rotating wheel and the measurement circuit.

FIG. 5 summarizes the results. The test was performed several times in each cage, and FIG. 5 shows the average value. From FIG. 5, the bearing according to each embodiment of the present invention, ie, L / X _A = 0, is compared with the conventional cage, ie, L / X _A = 0.
When X _A = 1 or more, it was found that the oil film forming ability was excellent.

[Bearing Vibration Deterioration Test] Performance comparison between the cage 1 in the above embodiment of the present invention and a conventional cage without the radial groove 4 was confirmed by a vibration deterioration test. The test bearing is a deep groove ball bearing 608 (inner diameter 8 mm x outer diameter 22 mm x
Width 7 mm). The ball 5 is a steel ball and has a diameter of 4 mm. Grease has a thickener of Li soap and a base oil of synthetic ester oil. The kinematic viscosity at 40 ° C. of the base oil is 33 mm ² / s. The ambient temperature was 80 ° C., and the axial load was 5 kgf. Test rotation speed is 1050mm
, And the inner ring was rotated, and the operation time was 200 h. The cage 1 is an injection-molded nylon crown type cage. The amount of vibration deterioration of the bearing before and after operation was measured, and the effect was confirmed.

FIG. 6 summarizes the results. From FIG. 6, it can be seen that the bearing according to each embodiment of the present invention has a small amount of vibration deterioration and an excellent lubrication life compared to the conventional cage, that is, the cage in which the radial groove is not provided.

[Torque Test] In the cage 1 according to the above embodiment of the present invention, the full depth dimension D (FIG. 3) is fixed, and the claw-opening end of the Hertzian contact trajectory on the outer diameter surface of the cage and the ball center are measured. The ratio (L / X _A ) of the half width dimension L of the radial groove 4 to the axial dimension X _{A of} FIG. Test bearings include deep groove ball bearings 608
(Inner diameter 8 mm x outer diameter 22 mm x width 7 mm) was used.
The ball 5 is a steel ball and has a diameter of 4 mm. Grease has a thickener of Li soap and a base oil of synthetic ester oil. The kinematic viscosity at 40 ° C. of the base oil is 33 mm ² / s. The ambient temperature was room temperature, and the axial load was 5 kgf. The test rotation speed was 5000 rpm, and the inner ring was rotated. The cage 1 is an injection-molded nylon crown type cage. The effect was confirmed by measuring the torque during the rotation of the bearing.

FIG. 7 summarizes the results. The test was performed several times in each cage, and FIG. 7 shows the average value. FIG. 7 shows that the bearing according to each embodiment of the present invention, ie, L / X _A = 0, compared to the conventional cage, ie, L / X _A = 0.
When X _A = 1 or more, the bearing incorporating the bearing according to each embodiment of the present invention has a low torque with respect to the conventional cage, that is, the cage in which the radial groove is not provided. The characteristics were shown.

[Oil Film Formation Test (2)] (High Speed Rotation) The performance of the cage 1 in the above embodiment of the present invention and the conventional cage without the radial groove 4 at the time of high speed rotation was confirmed by an oil film formation test. did. As a test bearing, a deep groove ball bearing 608 (inner diameter 8 mm × outer diameter 22 mm × width 7 mm) was used. The ball 5 is a steel ball and has a diameter of 4 mm. Grease has a thickener of urea, and the base oil is an ester synthetic oil. The kinematic viscosity at 40 ° C. of the base oil is 41 mm ² / s. The ambient temperature was 60 ° C., and the axial preload was 3 kgf. The test rotation speed was 20000 rpm, and the inner ring was rotated. The cage 1 is an injection molded nylon crown cage. A DC circuit for measuring the contact resistance between the stationary wheel (outer wheel) and the rotating wheel (inner wheel) during operation was constructed, and the oil film formation rate of the bearing was evaluated from the resistance value. In this case, the oil film formation rate refers to a voltage applied to the bearing of 1 V
The following equation is defined from the voltage drop V between the stationary wheel and the rotating wheel during the test. Oil film formation rate = 100 × V. Note that a mercury slip ring attached to the end of the rotating shaft was used for connection between the rotating wheel and the measurement circuit.

Table 1 summarizes the results. The test was performed several times in each cage, and Table 1 shows the average value. The width of the radial groove 4 of the cage according to each embodiment of the present invention in which the oil film formation rate was evaluated was constant. The axial section of the bottom surface of the radial groove 4 was arc-shaped. The groove depth in Table 1 indicates the maximum depth from the virtual spherical surface along the inner surface of the pocket 2 of the retainer 1 to the bottom surface of the radial groove 4. From Table 1, it was confirmed that the oil film forming ability of the bearing according to each embodiment of the present invention was superior to the oil film forming ability of the bearing incorporating the conventional cage.

[0039]

[Table 1]

[0040]

The ball bearing of the present invention has a radial groove penetrating from the inner diameter surface to the outer diameter surface of the cage, and the cross-sectional shape of the groove is arc-shaped. The state of formation of the film can be improved, and the lubrication life of the bearing can be extended. By providing the groove, the contact area between the ball and the inner surface of the pocket is reduced, so that the effect of reducing the torque of the bearing can be obtained.

[Brief description of the drawings]

FIG. 1 is a sectional view of a ball bearing according to an embodiment of the present invention.

FIG. 2 is an external explanatory view showing a state where a ball is held in a part of the cage.

FIG. 3 is a partial cross-sectional view along the axial direction at a PCD position of the cage.

FIG. 4 is a partial cross-sectional view of the retainer at a ball center position as viewed from an axial direction.

FIG. 5 is a graph showing a relationship between a dimensional ratio between a groove width and a Hertzian contact track width and an oil film formation rate.

FIG. 6 is a graph showing a relationship between a depth dimension of a groove and a vibration deterioration amount.

FIG. 7 is a graph showing a relationship between a dimensional ratio between a groove width and a Hertzian contact track width and torque.

[Explanation of symbols]

 1: Cage 2: Pocket 3: Claw 4: Radial groove 5: Ball 6: Inner ring 7: Outer ring 8: Seal

 ────────────────────────────────────────────────── ─── Continued on the front page F term (reference) 3J101 AA02 AA32 AA42 AA54 AA62 AA72 BA22 BA50 BA53 BA54 CA08 CA15 EA31 EA63 FA31 GA53

Claims (6)

[Claims] 1. A plurality of balls interposed between an inner ring and an outer ring are held in pockets provided in a ring-shaped retainer,
In a ball bearing in which the inner surface of the pocket is spherical, a radial groove penetrating from the inner surface to the outer surface of the cage is provided on the inner surface of the pocket of the cage, and the cross-sectional shape of the groove is arc-shaped. And ball bearings. 2. The ball bearing according to claim 1, wherein the axial position of the groove of the cage is a position including the center of the ball within the groove width. 3. In a state where an axial preload is applied between the inner ring and the outer ring, the contact point of the raceway surface of the inner and outer rings with the ball is located within the groove width of the radial groove. The ball bearing according to claim 2, wherein the position and width of the groove are set. 4. A deep groove ball bearing, wherein the larger one of the major axis radii a _A and a _B of each Hertz contact ellipse between the ball and the raceway surfaces of the inner ring and the outer ring calculated from the applied load and the bearing internal specifications is determined. a, the contact angle is θ, the outer diameter of the cage is Dco,
When the retainer inner diameter is Dci and the pitch circle diameter of the ball is Dp, the axial center of the groove is the axial center of the pocket surface of the retainer, and the axial width 2L of the groove is expressed by the following formula: Equation 1 2. The ball bearing according to claim 1, wherein the dimension is larger one of 2X _A and 2X _B defined by: 5. The ball bearing according to claim 1, wherein a vertical cross-sectional shape of the groove bottom surface of the groove of the cage is an arc concentric with a spherical inner surface of the pocket. 6. The ball bearing according to claim 1, wherein the cage is made of resin and has a crown shape.