JP7699528B2 - Sealed ball bearing - Google Patents

Description

This invention relates to a sealed ball bearing in which a seal member is provided between the inner and outer rings.

Ball bearings are often used to support rotating shafts in automobiles, industrial machinery, and other equipment. In general, a ball bearing has an inner ring, an outer ring arranged coaxially and radially outward of the inner ring, a number of balls arranged in the annular space between the inner ring and the outer ring, and a cage that holds the balls.

As an example of a cage, as disclosed in Patent Document 1, a cage (so-called crown cage) is known that has a cage annular portion that extends circumferentially adjacent to the ball passage area, and a cantilever-shaped cage claw portion that extends axially from the cage annular portion between adjacent balls in the circumferential direction. The cage claw portion has a ball guide surface that faces the surface of the ball, and this ball guide surface is a concave spherical surface that follows the surface of the ball so as to embrace the ball.

For example, as in Patent Document 2, sealed ball bearings are sometimes used in which the axial end openings of the annular space formed between the inner and outer rings are sealed with annular sealing members to prevent foreign matter from entering the ball bearing from the outside and to prevent lubricants (lubricating oil, grease, etc.) from leaking from the inside to the outside of the ball bearing.

Patent No. 3035766 International Publication No. WO 2016/143786

In recent years, in the field of electric vehicles such as EVs (battery electric vehicles) and HEVs (hybrid electric vehicles), efforts are being made to increase the speed of electric motors in order to reduce their size and weight. Ball bearings that support the rotating shaft to which the rotation of such electric motors is input are sometimes used under conditions where the dmn value (ball pitch circle diameter dm (mm) × rotation speed n (min ^-1 )) exceeds 2 million.

The inventors of this application have considered using a crown cage in ball bearings that support rotating shafts that rotate at high speeds in EVs, HEVs, and other vehicles.

However, it has been found that when a crown-shaped cage is used in a ball bearing that rotates at high speed, the centrifugal force acting on the cantilever-shaped cage claws causes twisting deformation in the cage ring that tilts the cage claws radially outward, and this deformation can cause the cage claws to interfere with the balls. If the cage claws interfere with the balls, this can cause the ball bearing to heat up.

In particular, when a bearing using a crown cage is also a sealed ball bearing, if the cage ring of the crown cage comes into contact with the seal member, there is a risk of abnormal heat generation due to the sliding resistance at the contact area, so it is necessary to reduce the axial width dimension of the cage ring so that the cage ring does not come into contact with the seal member. This makes it difficult to increase the rigidity of the cage ring, and the centrifugal force acting on the cage claws tends to cause torsional deformation in the cage ring, making it easy for the cage claws to interfere with the balls.

As such, when a crown cage was used in a sealed ball bearing, it was difficult to use the sealed ball bearing in high-speed rotation applications. Also, when the space where the bearing was to be installed was narrow and it was necessary to keep the width of the bearing small, it was difficult to prevent the crown cage from coming into contact with the seal member, so there were cases where it was necessary to abandon the use of a sealed ball bearing and use an open-type ball bearing with both axial ends open without a seal member.

The problem this invention aims to solve is to provide a sealed ball bearing suitable for use at high speeds.

In order to solve the above problems, the present invention employs the following structure in a sealed ball bearing.
With the inner circle,
an outer ring provided coaxially on the radially outer side of the inner ring;
A plurality of balls are mounted in an annular space formed between the inner ring and the outer ring;
an annular seal member provided at one end opening in the axial direction of the annular space;
a cage for holding the plurality of balls;
The cage has a cage ring portion extending in a circumferential direction through a region axially sandwiched between a ball passage region and the seal member, and a cantilever-shaped cage claw portion extending in the axial direction from the cage ring portion between adjacent balls in the circumferential direction,
the seal member has a seal-side sliding surface that faces the cage in the axial direction,
the cage has a cage-side sliding surface that faces the seal-side sliding surface in the axial direction,
a plurality of axial projections are formed at intervals in a circumferential direction on one of the seal-side sliding surface and the cage-side sliding surface;
Each of the axial projections has a sliding contact tip surface that extends linearly in the radial direction in a cross-sectional view perpendicular to the circumferential direction,
The other sliding surface is in sliding contact with the sliding tip surface during rotation of the bearing, and a smooth sliding plane is formed over the entire circumference,
A sealed ball bearing characterized in that the sliding tip surface and the sliding plane are arranged to face each other non-parallel and at an angle in a direction in which the distance between them increases from the radial inner side to the radial outer side when the bearing is stationary.

In this way, when the sliding tip surface of the axial protrusion formed on one of the sliding surfaces of the seal side sliding surface and the cage side sliding surface slides against the smooth sliding plane formed over the entire circumference on the other sliding surface during rotation of the bearing, the sliding tip surface of the axial protrusion has a shape that extends linearly in the radial direction, so the axial protrusion is less likely to push aside the lubricating oil and the lubricating oil is more likely to be drawn into the sliding portion of the axial protrusion. As a result, an oil film is formed between the sliding tip surface of the axial protrusion and the sliding plane due to the wedge film effect, and this oil film creates a fluid lubrication state between the sliding tip surface and the sliding plane, making it possible to keep the contact resistance between the cage and the seal member extremely small.

In addition, the sliding tip surface and the sliding plane are arranged to face each other non-parallel at an angle in the direction in which the distance widens from the radial inside to the radial outside when the bearing is stationary, so that when the bearing rotates at high speed, it is possible to stably form an oil film between the sliding tip surface and the sliding plane due to the wedge film effect. In other words, assuming a configuration in which the sliding tip surface and the sliding plane face each other in parallel when the bearing is stationary, when the bearing rotates at high speed, the cage ring portion is deformed by the centrifugal force acting on the cage claw portion, and this deformation tilts the cage side sliding surface, making the sliding tip surface and the sliding plane non-parallel, which creates the problem that it is difficult to form an oil film between the sliding tip surface and the sliding plane due to the wedge film effect. To address this issue, a configuration is adopted in which the sliding tip surface and the sliding plane face each other non-parallel and at an angle in the direction in which the distance between them increases from the radial inside to the radial outside when the bearing is stationary. When the bearing rotates at high speed, the centrifugal force acting on the cage claws causes the cage annular portion to deform, and this deformation brings the sliding tip surface and the sliding plane closer to parallel, making it possible to stably form an oil film between the sliding tip surface and the sliding plane due to the wedge film effect.

The angle between the sliding tip surface and the sliding plane when the bearing is stationary can be set in the range of 0.5° to 6°.

The axial length of the cage claw portion is set to be greater than the radius of the ball,
When the cage claw portion has a pocket side surface that faces the ball in the circumferential direction,
It is preferable that the portion of the pocket side surface that receives the ball circumferentially be configured to have a flat shape so that the pocket side surface does not interfere with the ball when the retainer claw portion moves radially outward due to centrifugal force.

In this way, when the bearing rotates at high speed, the centrifugal force acting on the cage claws causes the cage annular portion to deform, and when this deformation causes the cage claws to move radially outward, it is possible to prevent the pocket side of the cage claws from interfering with the balls and generating abnormal heat.

The pocket side surface can be a plane that runs along a straight line that passes through the bearing center and extends radially.

The pocket side surface can be a plane that runs along parallel straight lines that face each other in the circumferential direction, sandwiching a straight line that connects the bearing center and the circumferential center of the retainer claw portion.

In this way, even if the retainer claw portion moves radially along the straight line connecting the bearing center and the circumferential center of the retainer claw portion, the distance between the balls and the pocket side surface does not change. Therefore, when the bearing rotates at high speed, the centrifugal force acting on the retainer claw portion causes the retainer annular portion to deform, and even when this deformation causes the retainer claw portion to move radially outward, the retainer maintains stable retention of the balls.

the cage annular portion has a pocket bottom surface that faces the balls in the axial direction,
The pocket bottom surface has a shape that extends linearly in the radial direction in a cross-sectional view perpendicular to the circumferential direction,
A configuration can be adopted in which the pocket side surface and the pocket bottom surface are connected to each other in a concave arc shape when viewed in a radial direction.

In this way, the pocket side and bottom are connected in a concave arc shape, so the mass of the axial tip portion of the retainer claw can be kept small while ensuring a sufficient cross-sectional area of the axial root portion of the retainer claw. This makes it possible to effectively suppress deflection of the retainer claw due to centrifugal force acting on the retainer claw.

It is preferable to form a padding portion at the radially inner end of the bottom surface of the pocket that protrudes axially inward without coming into contact with the ball.

In this way, it is possible to effectively prevent the retainer annular portion from being damaged by the influence of centrifugal force. In other words, when the retainer annular portion is torsionally deformed by the centrifugal force acting on the retainer claw portions, stress concentration is likely to occur in the radially inner portion of the retainer annular portion at a position corresponding to the middle of the circumferentially adjacent retainer claw portions (the radially inner end of the pocket bottom surface). Therefore, by forming a padding portion that protrudes axially inward at the radially inner end of the pocket bottom surface, it is possible to effectively prevent the retainer annular portion from being damaged by stress concentration caused by centrifugal force.

an outer diameter side oil groove is formed on a radially outer surface of the retainer claw portion, the outer diameter side oil groove extending in an axial direction from a tip of the retainer claw portion toward the retainer annular portion;
an inner diameter side oil groove is formed on a radially inner surface of the retainer claw portion, the inner diameter side oil groove extending in an axial direction from a tip of the retainer claw portion toward the retainer annular portion;
It is preferable that the retainer claw portion is configured so that a cross-sectional shape perpendicular to the axial direction is an H-shape formed by the outer diameter side oil groove and the inner diameter side oil groove, the cross-sectional shape being open radially outward and radially inward.

In this way, the cross-sectional shape of the retainer claw portion is H-shaped due to the outer diameter side oil groove formed on the radial outer surface of the retainer claw portion and the inner diameter side oil groove formed on the radial inner surface of the retainer claw portion, so the mass of the retainer claw portion can be reduced while ensuring the second moment of area of the retainer claw portion (resistance of the retainer claw portion to deformation due to bending moment). This makes it possible to minimize the torsional deformation of the retainer annular portion and the bending deformation of the retainer claw portion itself caused by the centrifugal force received by the retainer claw portion when the bearing rotates at high speed.

the outer diameter side oil groove is formed such that a position of a groove bottom surface of the outer diameter side oil groove gradually changes radially outward from a tip side to a base side of the retainer claw portion,
It is preferable to adopt a configuration in which the cage annular portion is formed with a planar rising surface rising radially outward from the base of the cage claw portion.

In this way, the outer diameter oil groove is formed so that the position of the groove bottom surface gradually changes radially outward from the tip side to the base side of the retainer claw portion. Therefore, when the bearing rotates, the lubricating oil in the outer diameter oil groove moves from the tip side to the base side of the retainer claw portion due to the pumping effect. The lubricating oil that moves in the outer diameter oil groove from the tip side to the base side of the retainer claw portion is then returned axially inward by the rising surface that rises radially outward from the base of the retainer claw portion. By repeating this action, it becomes possible to efficiently circulate the lubricating oil in the bearing.

A configuration can be adopted in which an outer diameter through groove that axially penetrates the outer periphery of the retainer annular portion is formed at a circumferential position corresponding to the retainer claw portion.

In this way, the lubricating oil can easily pass through the outer diameter through groove between the area axially inside the retainer annular portion and the area axially outside the retainer annular portion, promoting circulation of the lubricating oil inside the bearing. In addition, since the outer diameter through groove is formed at a circumferential position of the retainer annular portion that corresponds to the retainer claw portion (i.e., a position of the retainer annular portion where rigidity is ensured by the retainer claw portion), it is possible to effectively prevent a decrease in rigidity of the retainer annular portion due to the formation of the outer diameter through groove.

Furthermore, it is preferable to adopt a configuration in which a non-penetrating axial groove is formed between the outer diameter side through grooves adjacent in the circumferential direction, extending in the axial direction around the outer periphery of the retainer annular portion, opening to the axially outer side of the retainer annular portion, and on the axially inner side of the retainer annular portion.

In this way, the lubricating oil can easily pass through the axial grooves and travel between the area axially inside the retainer annular portion and the area axially outside the retainer annular portion, promoting the circulation of the lubricating oil inside the bearing. In addition, since the axial grooves are formed so as not to penetrate the axial inside of the retainer annular portion, it is possible to effectively prevent damage to the retainer annular portion due to stress concentration. In other words, when the retainer annular portion is torsionally deformed by the centrifugal force acting on the retainer claw portion, stress concentration is likely to occur on the axial inside of the outer periphery of the retainer annular portion. Therefore, by forming the axial grooves on the outer periphery of the retainer annular portion so as not to penetrate the axial inside, it is possible to effectively prevent damage to the retainer annular portion due to stress concentration.

It is preferable that the axial length of the axial groove be set to 2/3 or less of the axial width of the retainer ring portion.

In this way, it is possible to effectively prevent a decrease in the rigidity of the retainer ring portion due to the formation of axial grooves on the outer periphery of the retainer ring portion, and to effectively prevent damage to the retainer ring portion due to stress concentration.

A configuration can be adopted in which an inner diameter through groove that penetrates the inner circumference of the retainer annular portion in the axial direction is formed at a circumferential position corresponding to the retainer claw portion.

In this way, the lubricating oil can easily pass through the inner diameter through groove between the area axially inside the retainer annular portion and the area axially outside the retainer annular portion, promoting circulation of the lubricating oil inside the bearing. In addition, since the inner diameter through groove is formed at a circumferential position of the retainer annular portion that corresponds to the retainer claw portion (i.e., a position of the retainer annular portion where rigidity is ensured by the retainer claw portion), it is possible to effectively prevent a decrease in rigidity of the retainer annular portion due to the formation of the inner diameter through groove.

the seal member has a seal-side inclined surface that extends linearly from the seal-side sliding surface toward the radially inner side and is inclined axially inward in a cross-sectional view perpendicular to the circumferential direction,
the retainer annular portion has, in a cross-sectional view perpendicular to the circumferential direction, a retainer-side inclined surface extending linearly inclined axially inward from the retainer-side sliding surface toward the radially inward direction, and a chamfered portion extending linearly inclined axially inward from the retainer-side sliding surface toward the radially outward direction,
The angle between the retainer-side inclined surface and the seal-side inclined surface is set to 10° or less,
It is preferable to adopt a configuration in which the angle formed between the chamfered portion and the seal-side sliding surface is set to be greater than 10° and equal to or less than 48°.

In this way, the angle between the retainer side inclined surface formed radially inward from the sliding portion of the retainer side sliding surface and the seal side inclined surface is smaller than the angle between the seal side sliding surface and the chamfered portion formed radially outward from the sliding portion of the retainer side sliding surface and the seal side sliding surface. In other words, the angle formed between the retainer and the seal member radially inward from the sliding portion of the retainer side sliding surface and the seal side sliding surface is smaller than the angle formed between the retainer and the seal member radially outward from the sliding portion of the retainer side sliding surface and the seal side sliding surface. Therefore, in the region sandwiched in the axial direction between the retainer and the seal member, a flow of lubricating oil occurs from the radial inner side of the sliding portion of the retainer side inclined surface and the seal side inclined surface to the radial outer side of the sliding portion of the retainer side inclined surface and the seal side inclined surface, making it possible to efficiently lubricate the inside of the bearing.

It is preferable to adopt a configuration in which a notch is formed at the tip of the retainer claw portion, which penetrates the groove shoulders on both sides of the outer diameter side oil groove and the inner diameter side oil groove in the circumferential direction.

This increases the amount of lubricating oil supplied to the balls from the circumferential direction, improving the lubrication performance of the balls.

The retainer may be made of resin.

The sealed ball bearing of the above configuration is particularly suitable for use as a bearing for an electric motor in an electric vehicle or as a bearing for an electric vehicle transmission that reduces the rotation of the electric motor.

When the sealed ball bearing of this invention rotates, the sliding tip surface of the axial protrusion formed on one of the seal side sliding surface and the cage side sliding surface slides against the smooth sliding plane formed over the entire circumference on the other sliding surface. Because the sliding tip surface of the axial protrusion has a shape that extends linearly in the radial direction, the axial protrusion is less likely to push aside the lubricating oil, and the lubricating oil is more likely to be drawn into the sliding portion of the axial protrusion. As a result, an oil film is formed between the sliding tip surface of the axial protrusion and the sliding plane due to the wedge film effect, and this oil film creates a fluid lubrication state between the sliding tip surface and the sliding plane, making it possible to keep the contact resistance between the cage and the seal member extremely small.

In addition, the sliding tip surface and the sliding plane are arranged to face each other non-parallel and at an angle in the direction in which the distance widens from the radial inside to the radial outside when the bearing is stationary, so that when the bearing rotates at high speed, an oil film can be stably formed between the sliding tip surface and the sliding plane due to the wedge film effect. In other words, when the bearing rotates at high speed, the cage annular portion is deformed by the centrifugal force acting on the cage claw portion, and this deformation causes the sliding tip surface and the sliding plane to approach parallelism, so that an oil film can be stably formed between the sliding tip surface and the sliding plane due to the wedge film effect.

FIG. 1 is a cross-sectional view showing a sealed ball bearing according to an embodiment of the present invention. 2 is a cross-sectional view of the sealed ball bearing of FIG. 1 taken along a cross section perpendicular to the axial direction. 2 is a cross-sectional view of the sealed ball bearing of FIG. 1 cut along the circumferential direction and viewed from the outer diameter side. 4 is a cross-sectional view taken along line IV-IV of FIG. A cross-sectional view taken along line V-V in FIG. FIG. 2 is a perspective view of the cage of FIG. 1 as seen from the cage claw portion side. 5A is an enlarged view of the vicinity of the retainer-side sliding surface and the seal-side sliding surface in FIG. 4 , and FIG. 5B is an enlarged view of the vicinity of the retainer-side sliding surface and the seal-side sliding surface when the sealed ball bearing in FIG. 4 is rotating at high speed. A cross-sectional view taken along line VIII-VIII in FIG. FIG. 3 is a diagram showing the relationship between the pocket side surface shown in FIG. 2 and the bearing center (center of the cage annular portion). FIG. 10 is a diagram showing a modification of the pocket side surface shown in FIG. 2, corresponding to FIG. 9; Schematic diagram of an electric vehicle transmission incorporating the sealed ball bearing of FIG. 1.

Figure 1 shows a sealed ball bearing 1 according to an embodiment of the present invention. This sealed ball bearing 1 has an inner ring 2, an outer ring 3 arranged coaxially on the radial outside of the inner ring 2, a number of balls 5 assembled at intervals in the circumferential direction in an annular space 4 formed between the inner ring 2 and the outer ring 3, a resin retainer 6 (hereinafter simply referred to as "retainer 6") that maintains the circumferential spacing of the number of balls 5, an annular seal member 7 that closes one of the end openings (the right end opening in the figure) of the end openings on both sides in the axial direction of the annular space 4, and an annular seal member 8 that closes the other end opening (the left end opening in the figure). A lubricant such as grease (not shown) is sealed in the annular space 4.

On the outer periphery of the inner ring 2, there are formed an inner ring raceway groove 9 with which the balls 5 roll and make contact, a pair of inner ring groove shoulders 10 located axially outward of the inner ring raceway groove 9, and a pair of sliding recesses 11 located axially outward of the inner ring groove shoulders 10. The inner ring raceway groove 9 is an arc groove with a concave arc-shaped cross section along the surface of the balls 5, and is formed extending circumferentially through the axial center of the outer periphery of the inner ring 2. The pair of inner ring groove shoulders 10 are bank-shaped parts extending circumferentially on both sides of the inner ring raceway groove 9 in the axial direction. The sliding recesses 11 are recesses extending circumferentially formed adjacent to the axial outside of the inner ring groove shoulders 10. The inner surfaces of the pair of sliding recesses 11 are in sliding contact with the seal members 7 and 8, respectively.

On the inner circumference of the outer ring 3, there are formed an outer ring raceway groove 12 with which the balls 5 roll and make contact, a pair of outer ring groove shoulders 13 located axially outward of the outer ring raceway groove 12, and a pair of seal fixing grooves 14 located axially outward of the outer ring groove shoulders 13. The outer ring raceway groove 12 is an arc groove with a concave arc-shaped cross section along the surface of the balls 5, and is formed extending circumferentially through the axial center of the inner circumference of the outer ring 3. The pair of outer ring groove shoulders 13 are bank-shaped parts extending circumferentially on both sides of the outer ring raceway groove 12 in the axial direction. The seal fixing groove 14 is a groove extending circumferentially formed adjacent to the axial outside of the outer ring groove shoulders 13. The seal members 7 and 8 are fitted and fixed in the seal fixing grooves 14.

The balls 5 are radially sandwiched between the outer ring raceway groove 12 and the inner ring raceway groove 9. This sealed ball bearing 1 is a deep groove ball bearing. That is, the axial width dimension of the outer ring raceway groove 12 is greater than half the diameter of the balls 5, and the axial width dimension of the inner ring raceway groove 9 is greater than half the diameter of the balls 5.

As shown in FIG. 4, the sealing member 7 is an annular member made up of an annular core metal 15 and a rubber material 16 (e.g., nitrile rubber, acrylic rubber, etc.) fixed to the core metal 15. The rubber material 16 is fixed to the core metal 15 by vulcanization insert molding. In other words, the rubber material 16 is vulcanized while the core metal 15 is placed in a mold for vulcanization molding the rubber material 16, so that the rubber material 16 is adhered and fixed to the surface of the core metal 15.

The seal member 7 has a fitting portion 17 that fits into the seal fixing groove 14, an annular plate portion 18 that extends radially inward from the fitting portion 17, and a seal lip 19 that slides against the inner surface of the sliding recess 11. The fitting portion 17 is provided at the outer diameter side end of the seal member 7. The seal lip 19 is provided at the inner diameter side end of the seal member 7. The surface of the inner surface of the sliding recess 11 against which the seal lip 19 slides is a cylindrical surface with a constant outer diameter along the axial direction.

The retainer 6 has a retainer annular portion 20 (see FIG. 5) that extends circumferentially through the area axially sandwiched between the ball 5 passage area and the seal member 7, and retainer claw portions 21 that extend axially from the retainer annular portion 20 between adjacent balls 5 in the circumferential direction. The retainer annular portion 20 and the retainer claw portions 21 are formed seamlessly from a resin composition. The resin composition forming the retainer annular portion 20 and the retainer claw portions 21 can be made of resin material only, but here a resin material with added fiber reinforcement is used.

Polyamide (PA) or super engineering plastics can be used as the base resin material for the resin composition. As polyamides, polyamide 46 (PA46), polyamide 66 (PA66), polynonamethylene terephthalamide (PA9T), etc. can be used. As super engineering plastics, polyether ether ketone (PEEK) and polyphenylene sulfide (PPS) can be used. As fiber reinforcing materials to be added to the resin material, glass fiber, carbon fiber, aramid fiber, etc. can be used.

The retainer claw portion 21 is formed in a cantilever shape with one axial end being a fixed end (the root of the retainer claw portion 21) fixed to the retainer ring portion 20 and the other axial end being a free end (the tip of the retainer claw portion 21). The axial length of the retainer claw portion 21 is set to be greater than the radius of the balls 5. The retainer claw portion 21 has a tapered shape with a radial thickness that gradually decreases from the side closer to the retainer ring portion 20 (the root side) to the side farther away (the tip side).

As shown in Fig. 2, the cage claw portion 21 has a pocket side surface 22 that faces the balls 5 in the circumferential direction. The portion of the pocket side surface 22 that receives the balls 5 in the circumferential direction is formed into a flat shape so that the pocket side surface 22 does not interfere with the balls 5 when the cage claw portion 21 moves radially outward due to centrifugal force. As shown in Fig. 9, when viewed from the axial direction, the pocket side surface 22 is a plane along a straight line _L1 that extends radially through the bearing center O (i.e., a plane that extends radially so that the bearing center O is located on an extension of the pocket side surface 22).

As shown in FIG. 3, the portion of the pocket side surface 22 that receives the balls 5 in the circumferential direction has a straight shape that extends straight in the axial direction without any circumferential inclination when viewed from the radial direction so that no axial component force is generated when the balls 5 are received. The cage ring portion 20 has a pocket bottom surface 23 that faces the balls 5 in the axial direction. The pocket side surface 22 and the pocket bottom surface 23 are connected in a concave arc shape when viewed from the radial direction.

As shown in FIG. 5, the pocket bottom surface 23 has a shape that extends linearly in the radial direction in a cross section perpendicular to the circumferential direction. A padding portion 24 that protrudes axially inward is integrally formed at the radially inner end of the pocket bottom surface 23. The padding portion 24 has a shape that protrudes radially inward to an extent that does not cause contact with the balls 5.

The cage 6 is an inner ring guide type cage that is positioned radially by contact with the inner ring 2. A cage guided surface 42 that is guided by the inner ring groove shoulder 10 on the outer circumference of the inner ring 2 is formed on the inner circumference of the cage annular portion 20. The cage guided surface 42 is an annular surface that is supported by sliding contact with the inner ring groove shoulder 10.

As shown in FIG. 4, an outer diameter oil groove 26 is formed on the radial outer surface 25 of the retainer claw portion 21, extending axially from the tip of the retainer claw portion 21 toward the retainer annular portion 20. An inner diameter oil groove 28 is formed on the radial inner surface 27 of the retainer claw portion 21, extending axially from the tip of the retainer claw portion 21 toward the retainer annular portion 20. The outer diameter oil groove 26 and the inner diameter oil groove 28 give the retainer claw portion 21 an H-shaped cross section perpendicular to the axial direction that is open radially outward and radially inward.

The outer diameter oil groove 26 is formed so that the position of the groove bottom surface of the outer diameter oil groove 26 gradually changes radially outward from the tip side to the root side of the retainer claw portion 21. The inner diameter oil groove 28 is formed so that the position of the groove bottom surface of the inner diameter oil groove 28 gradually changes radially inward from the tip side to the root side of the retainer claw portion 21. The retainer annular portion 20 is formed with a flat rising surface 29 that rises radially outward from the root of the retainer claw portion 21. The rising surface 29 is a plane perpendicular to the axial direction, and intersects with a step with the end of the root side of the retainer claw portion 21 on the inner surface of the outer diameter oil groove 26. In addition, as shown in FIG. 6, the rising surface 29 also intersects with a step with the end of the root side of the retainer claw portion 21 on the radial outer surface 25 of the retainer claw portion 21 (the surface on which the outer diameter oil groove 26 is not formed).

As shown in Figures 2 and 4, an outer diameter side through groove 30 that axially penetrates the outer circumference of the retainer annular portion 20 is formed at a circumferential position corresponding to the retainer claw portion 21 on the outer circumference of the retainer annular portion 20. Similarly, an inner diameter side through groove 31 that axially penetrates the inner circumference of the retainer annular portion 20 is formed at a circumferential position corresponding to the retainer claw portion 21 on the inner circumference of the retainer annular portion 20. The inner diameter side through groove 31 is in communication with the inner diameter side oil groove 28 of the retainer claw portion 21.

As shown in Figures 5 and 6, axial grooves 32 are formed on the outer periphery of the retainer annular portion 20 between the outer diameter side through grooves 30 adjacent in the circumferential direction, and extend in the axial direction around the outer periphery of the retainer annular portion 20. The axial grooves 32 are formed so as to be open on the axial outside of the retainer annular portion 20 and not to penetrate the axial inside of the retainer annular portion 20. In other words, the axial grooves 32 are formed only on the outer periphery of the retainer annular portion 20 from the axial middle portion to the axial outer end, and are not formed on the axial inner end of the outer periphery of the retainer annular portion 20. The axial length of the axial grooves 32 is set to 2/3 or less of the axial width of the retainer annular portion 20.

As shown in Figures 3 and 4, a notch 33 is formed at the tip of the retainer claw portion 21, circumferentially penetrating the groove shoulders on both sides of the outer diameter oil groove 26 and the inner diameter oil groove 28. By providing this notch 33, it is possible to increase the amount of lubricating oil supplied to the balls 5 from the circumferential direction, thereby improving the lubrication performance of the balls 5.

As shown in FIG. 4, the seal member 7 has a seal side sliding surface 34 that faces the retainer 6 in the axial direction. The seal side sliding surface 34 is an annular surface formed around the entire circumference on the axial inner side of the annular plate portion 18 of the seal member 7. The seal side sliding surface 34 is formed on the surface of the rubber material 16, not on the surface of the core metal 15. The retainer 6 also has a retainer side sliding surface 35 that faces the seal side sliding surface 34 in the axial direction. The retainer side sliding surface 35 is an annular surface formed around the entire circumference on the axial outer side of the retainer annular portion 20.

As shown in FIG. 3, a number of axial protrusions 36 are formed at regular intervals in the circumferential direction on the seal side sliding surface 34. In the figure, the axial height of the axial protrusions 36 is exaggerated to make the presence of the axial protrusions 36 easier to understand, but the axial height of the axial protrusions 36 is extremely small, at 0.5 mm or less.

As shown in FIG. 7(a), each axial protrusion 36 has a sliding tip surface 37 that extends linearly in the radial direction in a cross section perpendicular to the circumferential direction. That is, the axial protrusion 36 has a shape in which apexes of the same height extend continuously in the radial direction, and the sliding tip surface 37 is formed by the apexes. As shown in FIG. 8, the axial protrusion 36 is formed so that the cross section perpendicular to the radial direction is an axially convex arc shape. The axial protrusion 36 may also be formed so that the cross section perpendicular to the radial direction is an axially convex trapezoid shape.

On the other hand, as shown in Figures 7(a) and 8, a smooth sliding plane 38 is formed around the entire circumference of the cage side sliding surface 35. As shown in Figure 7(a), the sliding tip surface 37 and the sliding plane 38 are arranged to face each other non-parallel at an angle α in a direction in which the distance between them increases from the radial inside to the radial outside when the bearing is stationary. Here, the angle α between the sliding tip surface 37 and the sliding plane 38 when the bearing is stationary is set to a range of 0.5° to 6° (preferably 1° to 5°).

As shown in FIG. 4, the axial protrusion 36 is positioned at a position overlapping the pitch circle of the balls 5 (a virtual circle connecting the centers of multiple balls 5) or radially outward from that. Here, "the axial protrusion 36 is positioned at a position overlapping the pitch circle of the balls 5" means that the virtual cylindrical surface passing through the pitch circle of the balls 5 passes through the position of the axial protrusion 36, and "the axial protrusion 36 is positioned radially outward from the pitch circle of the balls 5" means that the entire axial protrusion 36 is positioned radially outward from the virtual cylindrical surface passing through the pitch circle of the balls 5. In the figure, the axial protrusion 36 is positioned radially outward from the pitch circle of the balls 5.

As shown in FIG. 7(a), the seal member 7 has a seal side inclined surface 39 that extends in a straight line inclined axially inward from the seal side sliding surface 34 toward the radial inside in a cross section perpendicular to the circumferential direction. The retainer annular portion 20 has a retainer side inclined surface 40 that extends in a straight line inclined axially inward from the retainer side sliding surface 35 toward the radial inside in a cross section perpendicular to the circumferential direction, and a chamfered portion 41 that extends in a straight line inclined axially inward from the retainer side sliding surface 35 toward the radial outside.

The retainer side inclined surface 40 and the seal side inclined surface 39 face each other in the axial direction at an angle β in a direction in which the distance between them increases from the radial outside to the radial inside. The angle β between the retainer side inclined surface 40 and the seal side inclined surface 39 is set to 10° or less. The chamfered portion 41 and the seal side sliding surface 34 face each other in the axial direction at an angle γ such that the distance between them increases gradually from the radial inside to the radial outside. The angle γ between the chamfered portion 41 and the seal side sliding surface 34 is set to be greater than 10° and less than 48°.

The above-mentioned sealed ball bearing 1 can be used as a bearing for an electric vehicle transmission 50 that reduces the rotation speed of an electric motor in an electric vehicle such as an EV (battery electric vehicle) or HEV (hybrid electric vehicle), as shown in Figure 11. The bearing for this electric vehicle transmission 50 rotates at a wide range of rotation speeds from low to high speeds while the vehicle is running, and when the bearing rotates at the highest speed, it is used under conditions where the dmn value (pitch circle diameter of balls 5 (mm) x rotation speed (min ^-1 )) exceeds 2 million.

The transmission 50 shown in FIG. 11 includes a stator 52 of an electric motor 51, a rotor 53 of the electric motor 51, a rotating shaft 54 connected to the rotor 53, a sealed ball bearing 1 that rotatably supports the rotating shaft 54, a second rotating shaft 55 and a third rotating shaft 56 arranged parallel to the rotating shaft 54, a first gear train 57 that transmits the rotation of the rotating shaft 54 to the second rotating shaft 55, and a second gear train 58 that transmits the rotation of the second rotating shaft 55 to the third rotating shaft 56. The stator 52 is an annular stationary member, and the rotor 53 as a rotating member is arranged inside the stator 52. When electricity is applied to the stator 52, the rotor 53 rotates due to the electromagnetic force acting between the stator 52 and the rotor 53, and the rotation of the rotor 53 is input to the rotating shaft 54.

In the sealed ball bearing 1 of this embodiment, as shown in Figs. 7(b) and 8, when the sliding tip surface 37 of the axial protrusion 36 formed on the seal side sliding surface 34 slides against the smooth sliding plane 38 formed on the retainer side sliding surface 35 during bearing rotation, the sliding tip surface 37 of the axial protrusion 36 has a shape that extends linearly in the radial direction (a shape in which the same height continues in the radial direction) as shown in Fig. 7(b), so that when the bearing rotates, the axial protrusion 36 does not easily push aside the lubricating oil, and the lubricating oil is easily drawn into the sliding portion of the axial protrusion 36. Therefore, as shown in Fig. 8, an oil film is formed between the sliding tip surface 37 of the axial protrusion 36 and the sliding plane 38 due to the wedge film effect, and the oil film creates a fluid lubrication state between the sliding tip surface 37 and the sliding plane 38, making it possible to keep the contact resistance between the retainer 6 and the seal member 7 extremely small. Therefore, it is possible to prevent abnormal heat generation due to the sliding resistance of the contact portion between the retainer 6 and the seal member 7.

Here, the lubrication state between the sliding surfaces is divided into boundary lubrication state and fluid lubrication state. The boundary lubrication state refers to a state in which the sliding surfaces are lubricated with an oil film consisting of several molecular layers (about 10 ^-5 to 10 ^-6 mm) of lubricating oil adsorbed on each sliding surface, and the fine unevenness of the sliding surfaces directly contact each other. On the other hand, the fluid lubrication state refers to a state in which an oil film (for example, about 10 ^-3 to 10 ^-1 mm) is formed between the sliding surfaces due to the wedge film effect, and the sliding surfaces do not directly contact each other due to the oil film (only indirect contact through the oil film occurs). When the wedge film effect occurs and the state of fluid lubrication is reached, the sliding resistance becomes almost zero.

As shown in FIG. 7(a), when the bearing is stationary, the sliding tip surface 37 and the sliding plane 38 are arranged to face each other non-parallel at an angle α in a direction in which the distance between them increases from the radial inside to the radial outside. This allows a stable oil film to be formed between the sliding tip surface 37 and the sliding plane 38 due to the wedge film effect when the bearing rotates at high speed.

In other words, if we assume that a configuration is adopted in which the sliding tip surface 37 and the sliding plane 38 shown in Figure 7 (a) are parallel and opposed to each other when the bearing is stationary, then when the bearing rotates at high speed, the centrifugal force acting on the retainer claw portion 21 shown in Figure 4 will deform the retainer annular portion 20, and this deformation will tilt the retainer side sliding surface 35, making the sliding tip surface 37 and the sliding plane 38 non-parallel, resulting in the problem that the sliding tip surface 37 will come into edge contact with the sliding plane 38 and making it difficult for an oil film to form between the sliding tip surface 37 and the sliding plane 38 due to the wedge film effect. To address this issue, the sealed ball bearing 1 of the above embodiment is configured so that when the bearing is stationary, the sliding tip surface 37 and the sliding plane 38 are opposed non-parallel at an angle α in the direction in which the distance between them increases from the radial inside to the radial outside, as shown in FIG. 7(a). Therefore, when the bearing rotates at high speed, the cage ring portion 20 is deformed by the centrifugal force acting on the cage claw portion 21 shown in FIG. 4, and this deformation causes the sliding tip surface 37 and the sliding plane 38 to approach parallelism as shown in FIG. 7(b), making it possible to stably form an oil film between the sliding tip surface 37 and the sliding plane 38 by the wedge film effect.

Furthermore, as shown in Figures 2 and 9, in this sealed ball bearing 1, the pocket side surface 22 is made flat along a straight line _L1 that extends radially through the bearing center O. Therefore, when the bearing rotates at high speed, the centrifugal force acting on the retainer claw portion 21 deforms the retainer annular portion 20, and when this deformation causes the retainer claw portion 21 to move radially outward, it is possible to prevent the pocket side surface 22 of the retainer claw portion 21 from interfering with the balls 5 and generating abnormal heat.

10, the pocket side surface 22 may be a plane along a parallel straight line _L3 that faces in the circumferential direction across a straight line _L2 that connects the bearing center O and the circumferential center of the retainer claw portion 21. Even in this case, when the retainer annular portion 20 is deformed by the centrifugal force acting on the retainer claw portion 21 and the retainer claw portion 21 moves radially outward due to the deformation, it is possible to prevent the pocket side surface 22 of the retainer claw portion 21 from interfering with the balls 5 (see FIG. 2) and generating abnormal heat.

10 is used, the distance between the balls 5 (see FIG. 2) and the pocket side surface 22 does not change when the retainer claw portions 21 move radially along the straight line _L2 connecting the bearing center O and the circumferential center of the retainer claw portions 21. Therefore, when the bearing rotates at high speed, the retainer annular portion 20 is deformed by the centrifugal force acting on the retainer claw portions 21, and even when the deformation causes the retainer claw portions 21 to move radially outward, the retention of the balls 5 by the retainer 6 is stable.

In addition, as shown in FIG. 3, in the sealed ball bearing 1 of the above embodiment, the pocket side surface 22 and the pocket bottom surface 23 are connected in a concave arc shape, so that the mass of the axial tip portion of the retainer claw portion 21 can be kept small while ensuring the cross-sectional area of the axial root portion of the retainer claw portion 21. Therefore, it is possible to effectively suppress the deflection of the retainer claw portion 21 due to the centrifugal force acting on the retainer claw portion 21.

In addition, as shown in Figs. 5 and 6, this sealed ball bearing 1 has a padding portion 24 formed at the radially inner end of the pocket bottom surface 23, so that it is possible to effectively prevent the retainer annular portion 20 from being damaged by the influence of centrifugal force. That is, when the retainer annular portion 20 is torsionally deformed by the centrifugal force acting on the retainer claw portion 21, stress concentration is likely to occur in the radially inner portion of the retainer annular portion 20 (the radially inner end of the pocket bottom surface 23) at a position corresponding to the middle of the retainer claw portions 21 adjacent in the circumferential direction. Therefore, by forming a padding portion 24 that protrudes axially inward at the radially inner end of the pocket bottom surface 23 as in the above embodiment, it is possible to effectively prevent the retainer annular portion 20 from being damaged by stress concentration caused by centrifugal force.

As shown in FIG. 4, the cross-sectional shape of the retainer claw portion 21 of this sealed ball bearing 1 is H-shaped due to the outer diameter side oil groove 26 formed on the radial outer surface 25 of the retainer claw portion 21 and the inner diameter side oil groove 28 formed on the radial inner surface 27 of the retainer claw portion 21, so that it is possible to reduce the mass of the retainer claw portion 21 while ensuring the second moment of area of the retainer claw portion 21 (resistance of the retainer claw portion 21 to deformation due to bending moment). Therefore, it is possible to reduce the torsional deformation of the retainer annular portion 20 and the bending deformation of the retainer claw portion 21 itself caused by the centrifugal force received by the retainer claw portion 21 when the bearing rotates at high speed.

As shown in FIG. 4, the outer diameter oil groove 26 of this sealed ball bearing 1 is formed so that the position of the bottom surface of the outer diameter oil groove 26 gradually changes radially outward from the tip side to the base side of the retainer claw portion 21. Therefore, when the bearing rotates, the lubricating oil in the outer diameter oil groove 26 moves from the tip side to the base side of the retainer claw portion 21 due to the pumping effect. The lubricating oil that moves in the outer diameter oil groove 26 from the tip side to the base side of the retainer claw portion 21 is returned axially inward by the rising surface 29 that rises radially outward from the base of the retainer claw portion 21. This action is repeated, making it possible to efficiently circulate the lubricating oil in the bearing.

As shown in FIG. 4, this sealed ball bearing 1 has an outer diameter side through groove 30 that axially penetrates the outer circumference of the retainer annular portion 20, and an inner diameter side through groove 31 that axially penetrates the inner circumference of the retainer annular portion 20. This allows the lubricating oil to easily pass through the outer diameter side through groove 30 and the inner diameter side through groove 31 between the area axially inside the retainer annular portion 20 and the area axially outside the retainer annular portion 20, promoting the circulation of the lubricating oil inside the bearing.

In addition, as shown in Figures 2 and 6, the outer diameter side through groove 30 and the inner diameter side through groove 31 are formed at circumferential positions in the retainer annular portion 20 that correspond to the retainer claw portions 21 (i.e., positions in the retainer annular portion 20 where rigidity is ensured by the retainer claw portions 21), so it is possible to effectively prevent a decrease in rigidity of the retainer annular portion 20 due to the formation of the outer diameter side through groove 30 and the inner diameter side through groove 31.

In addition, as shown in FIG. 5, this sealed ball bearing 1 has an axial groove 32 formed on the outer periphery of the retainer annular portion 20, which makes it easier for the lubricating oil to pass through the axial groove 32 and move between the area axially inside the retainer annular portion 20 and the area axially outside the retainer annular portion 20, thereby promoting the circulation of the lubricating oil inside the bearing.

As shown in FIG. 6, the axial grooves 32 are formed so as not to penetrate the axial inner side of the retainer annular portion 20, so that damage to the retainer annular portion 20 due to stress concentration can be effectively prevented. That is, when the retainer annular portion 20 is torsionally deformed by the centrifugal force acting on the retainer claw portion 21, stress concentration is likely to occur on the axial inner side of the outer periphery of the retainer annular portion 20. Therefore, by forming the axial grooves 32 on the outer periphery of the retainer annular portion 20 so as not to penetrate the axial inner side, damage to the retainer annular portion 20 due to stress concentration can be effectively prevented.

In addition, in this sealed ball bearing 1, the axial length of the axial groove 32 shown in FIG. 5 is set to 2/3 or less of the axial width of the retainer annular portion 20, so that the formation of the axial groove 32 on the outer periphery of the retainer annular portion 20 effectively prevents a decrease in the rigidity of the retainer annular portion 20, and effectively prevents damage to the retainer annular portion 20 due to stress concentration.

7(a), the angle β between the retainer side inclined surface 40 formed radially inward from the sliding portion of the retainer side sliding surface 35 and the seal side sliding surface 34 and the seal side inclined surface 39 is smaller than the angle γ between the chamfered portion 41 formed radially outward from the sliding portion of the retainer side sliding surface 35 and the seal side sliding surface 34 and the seal side sliding surface 34. In other words, the angle β formed between the retainer 6 and the seal member 7 radially inward from the sliding portion of the retainer side sliding surface 35 and the seal side sliding surface 34 is smaller than the angle γ formed between the retainer 6 and the seal member 7 radially outward from the sliding portion of the retainer side sliding surface 35 and the seal side sliding surface 34. Therefore, in the area sandwiched axially between the retainer 6 and the seal member 7, a flow of lubricating oil occurs from the radially inner side of the sliding contact area between the retainer side inclined surface 40 and the seal side inclined surface 39 toward the radially outer side of the sliding contact area between the retainer side inclined surface 40 and the seal side inclined surface 39, making it possible to efficiently lubricate the inside of the bearing.

In the above embodiment, an axial protrusion 36 having a sliding tip surface 37 is formed on the seal side sliding surface 34, and a sliding plane 38 against which the sliding tip surface 37 slides is formed on the cage side sliding surface 35. However, the configurations of the seal side sliding surface 34 and the cage side sliding surface 35 may be reversed. In other words, the cage side sliding surface 35 may be formed with an axial protrusion 36 having a sliding tip surface 37, and the seal side sliding surface 34 may be formed with a sliding plane 38 against which the sliding tip surface 37 slides.

In the above embodiment, the cage 6 is described as being made of a resin cage formed only from a resin composition. However, it is also possible to use a resin cage in which a metal ring-shaped core is insert-molded into the cage annular portion 20 when the cage annular portion 20 and the cage claw portion 21 are molded from a resin composition. It is also possible to use a soft steel cage in which the cage annular portion 20 and the cage claw portion 21 are integrally formed from soft steel.

The embodiments disclosed herein should be considered to be illustrative and not restrictive in all respects. The scope of the present invention is indicated by the claims, not the above description, and is intended to include all modifications within the meaning and scope of the claims.

Reference Signs List 1 Sealed ball bearing 2 Inner ring 3 Outer ring 4 Annular space 5 Ball 6 Cage 7 Seal member 20 Cage annular portion 21 Cage claw portion 22 Pocket side surface 23 Pocket bottom surface 24 Build-up portion 25 Radial outer surface 26 Outer diameter oil groove 27 Radial inner surface 28 Inner diameter oil groove 29 Rising surface 30 Outer diameter through groove 31 Inner diameter through groove 32 Axial groove 33 Notch 34 Seal side sliding surface 35 Cage side sliding surface 36 Axial protrusion 37 Sliding tip surface 38 Sliding flat surface 39 Seal side inclined surface 40 Cage side inclined surface 41 Chamfered portion 50 Electric vehicle transmission 51 Electric motor α, β, γ Angles L ₁ , L ₂ , L ₃ Straight line O Bearing center

Claims (17)

Inner circle (2) and
an outer ring (3) provided coaxially on the radially outer side of the inner ring (2);
A plurality of balls (5) are assembled in an annular space (4) formed between the inner ring (2) and the outer ring (3);
an annular seal member (7) provided at one axial end opening of the annular space (4);
A cage (6) for holding the plurality of balls (5),
The retainer (6) has a retainer annular portion (20) extending in the circumferential direction in a region sandwiched in the axial direction between a passage area of the balls (5) and the seal member (7), and a cantilever-shaped retainer claw portion (21) extending in the axial direction from the retainer annular portion (20) between adjacent balls (5) in the circumferential direction,
The seal member (7) has a seal-side sliding surface (34) that faces the cage (6) in the axial direction,
The cage (6) has a cage-side sliding surface (35) that faces the seal-side sliding surface (34) in the axial direction,
One of the seal-side sliding surface (34) and the cage-side sliding surface (35) has a plurality of axial projections (36) formed at intervals in the circumferential direction,
Each of the axial projections (36) has a sliding contact tip surface (37) that extends linearly in the radial direction in a cross-sectional view perpendicular to the circumferential direction,
The other sliding surface is in sliding contact with the sliding tip surface (37) during rotation of the bearing, and a smooth sliding plane (38) is formed over the entire circumference.
The sliding tip surface (37) and the sliding plane (38) are arranged to face each other non-parallel at an angle (α) in a direction in which the distance between them increases from the radial inner side toward the radial outer side when the bearing is stationary. The sealed ball bearing according to claim 1, wherein the angle (α) between the sliding tip surface (37) and the sliding plane (38) when the bearing is stationary is set in the range of 0.5° to 6°. The axial length of the retainer claw portion (21) is set to be greater than the radius of the ball (5),
The cage claw portion (21) has a pocket side surface (22) that faces the ball (5) in the circumferential direction,
3. A sealed ball bearing as described in claim 1 or 2, wherein the portion of the pocket side surface (22) that receives the ball (5) in the circumferential direction is made flat so that the pocket side surface (22) does not interfere with the ball (5) when the retainer claw portion (21) moves radially outward due to centrifugal force. 4. The sealed ball bearing according to claim 3, wherein the pocket side surface (22) is a flat surface along a straight line (L ₁ ) extending radially through the bearing center (O). The sealed ball bearing according to claim 3, wherein the pocket side surface (22) is a plane extending along parallel straight lines ( _L3 ) that face each other in the circumferential direction with respect to a straight line ( _L2 ) that connects the bearing center (O) and the circumferential center of the retainer claw portion (21). The cage annular portion (20) has a pocket bottom surface (23) that faces the balls (5) in the axial direction,
The pocket bottom surface (23) has a shape that extends linearly in the radial direction in a cross-sectional view perpendicular to the circumferential direction,
6. The sealed ball bearing according to claim 3, wherein the pocket side surface (22) and the pocket bottom surface (23) are connected in a concave arc shape when viewed in a radial direction. A sealed ball bearing as described in claim 6, in which a padding portion (24) that protrudes axially inward without contacting the ball (5) is formed at the radially inner end of the pocket bottom surface (23). An outer diameter side oil groove (26) is formed on a radial outer surface (25) of the retainer claw portion (21) and extends in the axial direction from a tip of the retainer claw portion (21) toward the retainer annular portion (20),
an inner diameter side oil groove (28) is formed on a radially inner surface (27) of the retainer claw portion (21) and extends in the axial direction from a tip of the retainer claw portion (21) toward the retainer annular portion (20);
8. A sealed ball bearing as claimed in any one of claims 1 to 7, wherein the retainer claw portion (21) has an H-shaped cross-sectional shape perpendicular to the axial direction, the cross-sectional shape being formed by the outer diameter side oil groove (26) and the inner diameter side oil groove (28) such that the cross-sectional shape is open radially outward and radially inward. the outer diameter side oil groove (26) is formed such that the position of the groove bottom surface of the outer diameter side oil groove (26) gradually changes radially outward from the tip side toward the base side of the retainer claw portion (21),
9. The sealed ball bearing according to claim 8, wherein the retainer annular portion (20) is formed with a flat rising surface (29) rising radially outward from the base of the retainer claw portion (21). A sealed ball bearing according to any one of claims 1 to 9, in which an outer diameter through groove (30) that axially penetrates the outer periphery of the retainer annular portion (20) is formed at a circumferential position corresponding to the retainer claw portion (21). A sealed ball bearing according to claim 10, in which a non-through axial groove (32) is formed between the outer diameter side through grooves (30) adjacent in the circumferential direction, the non-through axial groove (32) extends axially around the outer periphery of the retainer annular portion (20), is open to the axial outside of the retainer annular portion (20), and is formed on the axial inside of the retainer annular portion (20). The sealed ball bearing according to claim 11, wherein the axial length of the axial groove (32) is set to 2/3 or less of the axial width of the retainer annular portion (20). A sealed ball bearing according to any one of claims 1 to 12, in which an inner diameter through groove (31) that axially penetrates the inner circumference of the retainer annular portion (20) is formed at a circumferential position corresponding to the retainer claw portion (21). The seal member (7) has a seal-side inclined surface (39) that extends linearly from the seal-side sliding surface (34) toward the radially inner side and is inclined axially inward in a cross-sectional view perpendicular to the circumferential direction,
The retainer annular portion (20) has a retainer-side inclined surface (40) extending linearly from the retainer-side sliding surface (35) toward the radially inward direction and inclined axially inward in a cross-sectional view perpendicular to the circumferential direction, and a chamfered portion (41) extending linearly from the retainer-side sliding surface (35) toward the radially outward direction and inclined axially inward,
The angle (β) between the retainer side inclined surface (40) and the seal side inclined surface (39) is set to 10° or less,
14. A sealed ball bearing according to claim 1, wherein the angle (γ) between the chamfered portion (41) and the seal-side sliding surface (34) is set to be greater than 10° and equal to or less than 48°. 10. A sealed ball bearing as claimed in claim 8 or 9, wherein a notch (33) is formed at the tip of the retainer claw portion (21) so as to circumferentially penetrate groove shoulders on both circumferential sides of the outer diameter side oil groove (26) and the inner diameter side oil groove (28 ). A sealed ball bearing according to any one of claims 1 to 15, wherein the retainer (6) is a resin retainer. A sealed ball bearing according to any one of claims 1 to 16, used as a bearing for an electric motor (51) of an electric vehicle or a bearing for an electric vehicle transmission (50) that reduces the rotation of the electric motor (51).

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