Ball bearing

Abstract

According to the present invention, the axial thickness a (mm) of an inter-pocket region (17) of a retainer ring (11) satisfies a < c with respect to the axial distance c (mm) from the lateral surface (19) on one side of the retainer ring (11) in the axial direction to the center O of a ball (5), and provided that dm (mm) is the pitch circle diameter of balls (5), Z (number of entities) is the number of balls (5), and Da (mm) is the diameter of a ball (5), the ball-to-ball circumference ratio r, defined by r = (dm × π / Z) / Da, is set to a magnitude between 2.6 and 11.4 inclusive.

Description

ball bearing

[0001] The present invention relates to a ball bearing.

[0002] BACKGROUND ART As bearings used in high speed rotation ranges, such as bearings that support the rotating shaft of an electric motor in an electric vehicle, ball bearings such as those disclosed in Patent Documents 1 and 2 are known.

[0003] The ball bearings in Patent Documents 1 and 2 include an outer ring with an outer ring raceway surface formed on its inner periphery, an inner ring with an inner ring raceway surface formed on its outer periphery, multiple balls in rolling contact with the outer ring raceway surface and the inner ring raceway surface, and a cage that holds the multiple balls. The cage has a cage annular portion extending circumferentially on one axial side of the balls, and multiple pairs of cage claws extending from the cage annular portion to the other axial side. Pockets are formed between each pair of cage claws to accommodate each ball. The multiple balls are arranged at intervals circumferentially in an annular bearing space formed between the inner periphery of the outer ring and the outer periphery of the inner ring.

[0004] The ball bearings disclosed in Patent Documents 1 and 2 are used under oil lubrication. That is, in the ball bearings disclosed in Patent Documents 1 and 2, the bearing space between the outer ring and the inner ring is not sealed with a seal member or the like, but is open to the outside of the bearing. During rotation of the bearing, lubricating oil supplied from outside the bearing is introduced into the bearing space, and the bearing rotates while lubricating the interior of the bearing with the lubricating oil.

[0005] Patent No. 5703894 Patent No. 4626183

[0006] The above ball bearing is measured by the dmn value (ball pitch diameter dm (mm) x rotation speed n (min －１ When used in a high-speed rotation range where the RH (relative humidity) is 650,000 or more, the balls and cage rotate at high speed, causing air in the bearing space between the outer and inner rings to move circumferentially at high speed, creating an air curtain. This air curtain makes it difficult for lubricating oil supplied from outside the bearing to enter the bearing space, which can lead to starvation (depletion of lubricating oil inside the bearing).

[0007] Here, in the case of a ball bearing that supports the main spindle of a machine tool, etc., it is possible to prevent starvation by increasing the amount of lubricating oil supplied to the ball bearing. However, in the case of a ball bearing that supports the rotating shaft of an electric motor used to drive an electric vehicle, or a ball bearing that supports the rotating shaft of a unit (e-Axle) that integrates an electric motor and a reducer, it is difficult to increase the amount of lubricating oil supplied to the ball bearing.

[0008] The problem to be solved by this invention is to provide a ball bearing that is less likely to cause starvation when used in a high speed rotation range.

[0009] The present inventors have noticed that when a ball bearing is used at high rotational speeds, the area where lubricating oil is particularly likely to be insufficient is the contact area between the balls and the inner ring raceway. Specifically, when a ball bearing rotates at high speeds, the lubricating oil supplied to the bearing space from outside the bearing tends to move radially outward due to centrifugal force. Meanwhile, the contact area between the balls and the inner ring raceway is located at the innermost radial position within the bearing space, making it more difficult for lubricating oil supplied from outside the bearing to reach this area compared to other areas. Therefore, the inventors have noticed that when a ball bearing is used at high rotational speeds, the area where lubricating oil is particularly likely to be insufficient is the contact area between the balls and the inner ring raceway. Therefore, if this contact area between the balls and the inner ring raceway can be effectively lubricated, it will be possible to effectively prevent starvation at high rotational speeds.

[0010] Based on this focus, the present invention provides a ball bearing having the following configuration in order to solve the above-mentioned problems. [Configuration 1] A ball bearing comprising: an outer ring having an outer ring raceway surface formed on its inner circumference; an inner ring having an inner ring raceway surface formed on its outer circumference; a plurality of balls mounted at intervals in an annular bearing space formed between the outer ring and the inner ring and in rolling contact with the outer ring raceway surface and the inner ring raceway surface; and a cage for holding the plurality of balls, wherein the bearing space is in communication with the outside of the bearing so as to introduce lubricating oil from outside the bearing, and the cage has a cage annular portion extending circumferentially on one axial side of the plurality of balls, and a plurality of pairs of cage claw portions extending from the cage annular portion on the other axial side, and wherein pockets for accommodating each of the balls are formed between each pair of cage claw portions, wherein an axial thickness a (mm) of a portion connecting the cage claw portions between circumferentially adjacent pockets of the cage annular portion satisfies a < c, where c (mm) is an axial distance from one axial side of the cage annular portion to the center of the ball, A ball bearing characterized in that, when the pitch circle diameter of the balls is dm (mm), the number of the balls is Z (pieces), and the diameter of the balls is Da (mm), the ball-to-ball circumferential length ratio r, defined as r = (dm × π ÷ Z) / Da, is set to a value of 2.6 or more and 11.4 or less.

[0011] With this configuration, the circumferential spacing between adjacent balls is long, and the axial thickness of the portion connecting the cage claws between adjacent pockets in the cage annular portion is thin, so when viewed from the outside in the radial direction, only a small portion of the inner ring raceway is covered by the cage, making it easier for lubricating oil supplied from outside the bearing to reach the contact area between the balls and the inner ring raceway.

[0012] Furthermore, because the circumferential spacing between adjacent balls is long, the time interval between the balls passing the same position on the inner ring raceway surface is long. This reduces the likelihood of insufficient lubrication at the contact point between the ball and the inner ring raceway surface. Specifically, as the balls roll along the inner ring raceway surface, they push aside the lubricant present on the inner ring raceway surface. If the next ball rolls along the same position on the inner ring raceway surface before the amount of lubricant present on the inner ring raceway surface has recovered, there is a risk of direct contact between the ball and the inner ring raceway surface without any intervening lubricant. To address this issue, increasing the circumferential spacing between adjacent balls increases the time interval between the balls passing the same position on the inner ring raceway surface. This makes it easier for the amount of lubricant present on the inner ring raceway surface to recover between the time the ball rolls along the inner ring raceway surface and pushes aside the lubricant present on the inner ring raceway surface, and the time the next ball rolls along the same position on the inner ring raceway surface. This reduces the likelihood of direct contact between the ball and the inner ring raceway surface without any intervening lubricant, reducing the likelihood of insufficient lubrication at the contact point between the ball and the inner ring raceway surface.

[0013] Furthermore, because the circumferential spacing between adjacent balls is long, there is a large gap inside the bearing through which lubricating oil can pass in the axial direction (a gap through which lubricating oil supplied from outside the bearing can pass axially through the bearing space between the outer ring and inner ring, avoiding the balls and cage).As a result, even when the ball bearing is used in a high-speed rotation range, a large amount of lubricating oil passes through the bearing space, making it easy to ensure a sufficient amount of lubricating oil to lubricate the contact area between the balls and the inner ring raceway surface.

[0014] Due to the synergistic effect of each of the above-mentioned actions, it becomes possible to effectively lubricate the contact area between the balls and the inner ring raceway surface with lubricating oil supplied from outside the bearing, and it becomes possible to effectively prevent starvation in the high-speed rotation range.

[0015] Furthermore, with the above configuration, the number of balls is small, and therefore the number of cage claws is also small. Therefore, when the ball bearing is used at high rotational speeds, it is possible to prevent the cage from being deformed by centrifugal force, causing each pair of cage claws to tilt radially outward. That is, when the cage rotates at high speed, centrifugal force acts on each pair of cage claws extending axially from the cage annular portion. The force transmitted from these cage claws causes torsional deformation of the cage annular portion, resulting in the entire cage being deformed so that each pair of cage claws tilt radially outward. If this deformation of the cage becomes too great, the cage claws may interfere with the balls, potentially resulting in abnormal heat generation. To address this problem, increasing the circumferential spacing between circumferentially adjacent balls reduces the number of balls and the number of cage claws, which reduces the total centrifugal force acting on the multiple pairs of cage claws when the cage rotates at high speed and also reduces the torsional deformation of the cage annular portion caused by the force transmitted from those cage claws.As a result, when a ball bearing is used in the high-speed rotation range, it is possible to prevent the cage from being deformed by centrifugal force so that each of the multiple pairs of cage claws tilts radially outward.

[0016] [Configuration 2] A ball bearing according to Configuration 1, wherein the axial thickness a (mm) of a portion of the retainer annular portion that connects the retainer claw portions between circumferentially adjacent pockets satisfies a > b, where b (mm) is the axial thickness of a portion of the retainer annular portion that corresponds to the bottom of the pocket.

[0017] By adopting this configuration, the strength of the annular portion of the cage can be ensured.

[0018] [Configuration 3] A ball bearing according to Configuration 1 or 2, wherein a cylindrical inner ring shoulder surface is formed on the outer periphery of the inner ring, axially adjacent to the inner ring raceway surface and radially facing the inner periphery of the retainer annular portion, and an annular oil-passing gap for allowing lubricating oil to pass is formed between the inner periphery of the retainer annular portion and the inner ring shoulder surface.

[0019] When this configuration is adopted, an oil gap is formed in a position close to the inner ring raceway surface (between the inner circumference of the retainer annular portion and the inner ring shoulder surface), making it possible to effectively lubricate the contact area between the balls and the inner ring raceway surface.

[0020] [Configuration 4] The ball bearing according to any one of Configurations 1 to 3, wherein a side surface on the other axial side of a portion connecting the retainer claw portions between the pockets adjacent to each other in the circumferential direction of the retainer annular portion is a plane perpendicular to the axial direction.

[0021] When this configuration is adopted, since the side surface on the other axial side between the pockets adjacent to each other in the circumferential direction of the retainer annular portion is a plane perpendicular to the axial direction, the air existing in the space between the balls adjacent to each other in the circumferential direction is not easily agitated by the circumferential movement of the retainer annular portion. Therefore, the lubricating oil existing in the bearing space easily reaches the contact portion between the ball and the inner ring raceway surface.

[0022] [Configuration 5] The ball bearing according to any one of Configurations 1 to 4, wherein an oil passage groove extending in the radial direction is formed on the inner surface of the pocket at an axial position corresponding to the center of the ball.

[0023] When this configuration is adopted, it becomes possible to introduce lubricating oil into the oil passage groove on the inner surface of the pocket of the retainer and effectively lubricate the contact portion between the retainer and the ball with the lubricating oil.

[0024] [Configuration 6] The ball bearing according to any one of Configurations 1 to 5, wherein an axial thickness b (mm) of a portion corresponding to the bottom of the pocket of the retainer annular portion is set to a size of 1 / 62 or more and 1 / 26 or less of the pitch circle diameter dm (mm) of the ball.

[0025] [Configuration 7] The ball bearing according to any one of Configurations 1 to 6, wherein an axial thickness a (mm) of a portion connecting the retainer claw portions between the pockets adjacent to each other in the circumferential direction of the retainer annular portion is set to a size of 1 / 70 or more and 1 / 30 or less of the pitch circle diameter dm (mm) of the ball.

[0026] [Configuration 8] The ball bearing according to any one of Configurations 1 to 7, wherein the retainer is formed of a resin composition in which a fiber reinforcing material is added to any one of a polyamide resin, a polyether ether ketone resin, and a polyphenylene sulfide resin.

[0027] [Configuration 9] The ball bearing according to any one of Configurations 1 to 7, wherein the retainer is formed of soft steel.

[0028] [Configuration 10] The ball bearing of each of the above configurations is suitable for use as a bearing for supporting the rotating shaft of an electric motor for driving an electric vehicle, or as a bearing for supporting the rotating shaft of a reducer for reducing the rotation of the electric motor.

[0029] That is, although bearings that support the rotating shaft of an electric motor used to drive an electric vehicle, or bearings that support the rotating shaft of a reducer that slows down the rotation of that electric motor, rotate at extremely high speeds, the magnitude of the load that the bearing supports is very small (for example, a few percent or less) compared to the basic dynamic load rating of a standard ball bearing, so it is possible to set the number of balls to a small number. For this reason, the use of the ball bearing according to the present invention is particularly suitable.

[0030] The ball bearing of this invention can effectively lubricate the contact area between the balls and the inner ring raceway surface with lubricating oil supplied from outside the bearing, making it possible to effectively prevent starvation in the high-speed rotation range.

[0031] 1 is a cross-sectional view taken along line II-II in FIG. 1; FIG. 2 is a partial cross-sectional view of the vicinity of the pocket of the cage shown in FIG. 2, viewed from the radially outer side; FIG. 1 is a perspective view of the cage shown in FIG. 1; FIG. 1 is a schematic view of an example of use in which the ball bearing shown in FIG. 1 is incorporated into a drive device for driving an electric vehicle; and FIG. 2 is a view of an example of a standard ball bearing.

[0032] Figure 1 shows a ball bearing 1 according to an embodiment of the present invention. This ball bearing 1 has an outer ring 2, an inner ring 3 provided coaxially and radially inward of the outer ring 2, a plurality of balls 5 mounted at regular intervals in the circumferential direction in an annular bearing space 4 (see Figure 2) formed between the outer ring 2 and the inner ring 3, and a cage 6 that holds the plurality of balls 5.

[0033] The axial direction is the direction parallel to the central axis of the outer ring 2 (the central axis of the bearing), the radial direction is the direction perpendicular to the central axis of the outer ring 2, and the circumferential direction is the direction along the circumference that revolves around the central axis of the outer ring 2.

[0034] As shown in Figure 2, the inner circumference of the outer ring 2 is formed with an outer ring raceway 7 with which the balls 5 roll and make contact, and a pair of outer ring shoulder surfaces 8 extending circumferentially and adjacent in the axial direction on either side of the outer ring raceway surface 7. The outer ring raceway surface 7 is the inner surface of a groove with an arc-shaped cross section that extends circumferentially around the inner circumference of the outer ring 2, and the outer ring shoulder surface 8 is a cylindrical surface with a constant inner diameter along the axial direction. The outer diameter of the outer ring 2 is set to be equal to or greater than 60 mm and equal to or less than 110 mm.

[0035] The outer periphery of the inner ring 3 is formed with an inner ring raceway 9 with which the balls 5 roll and make contact, and a pair of inner ring shoulder surfaces 10 extending circumferentially and adjacent in the axial direction on both sides of the inner ring raceway 9. The inner ring raceway 9 is the outer surface of a groove with an arc-shaped cross section that extends circumferentially around the outer periphery of the inner ring 3, and the inner ring shoulder surface 10 is a cylindrical surface with a constant outer diameter along the axial direction. The inner diameter of the inner ring 3 is set to be equal to or greater than 35 mm and equal to or less than 80 mm.

[0036] The balls 5 are in rolling contact with an outer ring raceway surface 7 and an inner ring raceway surface 9. The outer ring raceway surface 7 is formed symmetrically with respect to the axial center of the outer ring 2, and the inner ring raceway surface 9 is also formed symmetrically with respect to the axial center of the inner ring 3. This ball bearing 1 is a deep groove ball bearing.

[0037] The bearing space 4 between the outer ring 2 and the inner ring 3 is not sealed on either axial side with a seal member or the like, and is in communication with the outside of the bearing. In other words, the bearing space 4 between the outer ring 2 and the inner ring 3 is in communication with the external space on both axial sides, so that lubricating oil supplied from outside the bearing can be introduced into the bearing space 4 between the outer ring 2 and the inner ring 3 while the bearing is rotating, and this lubricating oil can lubricate the inside of the bearing.

[0038] The cage 6 has a cage annular portion 11 extending circumferentially on one axial side of the balls 5 (the right side in the figure), and multiple cage claws 12 extending from the cage annular portion 11 to the other axial side (the left side in the figure) between circumferentially adjacent balls 5. Each cage claw 12 is formed in a cantilever shape with its end on one axial side (the right side in the figure) fixed to the cage annular portion 11 as a fixed end and its end on the other axial side (the left side in the figure) as a free end. The cage annular portion 11 and each cage claw 12 are formed seamlessly and integrally from a resin composition made by adding fiber reinforcing material to a resin material.

[0039] The resin material constituting the resin composition may be polyamide resin (PA), polyether ether ketone resin (PEEK), or polyphenylene sulfide resin (PPS). Examples of polyamide resin (PA) that may be used include polyamide 46 (PA46), polyamide 66 (PA66), and polynonamethylene terephthalamide (PA9T). Examples of fiber reinforcement added to the resin material may include glass fiber, carbon fiber, and aramid fiber. The fiber reinforcement is blended in at a ratio of 10 to 50% by weight of the resin composition forming the cage 6.

[0040] The inner periphery of the cage annular portion 11 faces radially the inner ring shoulder surface 10 on one axial side (the right side in the figure). An annular oil passage gap 13 is formed between the inner periphery of the cage annular portion 11 and the inner ring shoulder surface 10, allowing lubricating oil supplied from outside the bearing to pass through. The radial width of the oil passage gap 13 is set to be 5% or more of the diameter of the balls 5. A second annular oil passage gap 14, allowing lubricating oil to pass through, is also formed between the outer periphery of the cage annular portion 11 and the outer ring shoulder surface 8 on one axial side (the right side in the figure).

[0041] As shown in FIG. 1 , multiple pairs of cage claws 12 (six pairs in the figure) are provided to correspond to the multiple balls 5. The number of pairs of cage claws 12 is the same as the number of balls 5 (six in the figure). The center-to-center distance between adjacent balls 5 in the circumferential direction is set to be larger than the center-to-center distance between adjacent balls 5 in the circumferential direction of a standard ball bearing (a general ball bearing having dimensions specified in Japanese Industrial Standard JIS B1512-1 "Rolling Bearings - Boundary Dimensions - Part 1: Radial Bearings"). Specifically, when the pitch circle diameter of the balls 5 is dm (mm), the number of balls 5 is Z (pieces), and the diameter of the balls 5 is Da (mm), the ball-to-ball circumferential length ratio r, defined as r = (dm × π ÷ Z) / Da, is set to be 2.6 or more (preferably 4.3 or more, more preferably 5.7 or more) and 11.4 or less. The pitch circle diameter dm (mm) of the balls 5 is the diameter of an imaginary circle connecting the centers O of the balls 5. The pitch circle diameter dm (mm) is equal to the midpoint between the outer diameter of the outer ring 2 and the inner diameter of the inner ring 3. The ball-to-ball circumferential length ratio r of a standard ball bearing such as that shown in Figure 6 is approximately 1.8. If this ratio r is less than 2.6, it becomes difficult to ensure lubrication during high-speed rotation, and if the ratio r is greater than 11.4, the circumferential spacing of the balls 5 will be too wide, creating a risk of the bearing disassembly.

[0042] For example, in the case of a standard ball bearing 1 in which the outer diameter of the outer ring 2 is 90 mm and the inner diameter of the inner ring 3 is 65 mm (specifically, in the case of a ball bearing with a bearing number of 6913, in this case the pitch circle diameter dm = 77.5 (mm) and the diameter of the balls 5 Da = 9 / 32 inch = 7.14375 (mm)), the number Z of balls 5 is 19. When the present invention is applied to a ball bearing 1 having the same dimensions, the number Z of balls 5 can be set to 13 or less (r = 2.6 or more), preferably 8 or less (r = 4.3 or more), and more preferably 6 or less (r = 5.7 or more).

[0043] Similarly, in the case of a standard ball bearing 1 in which the outer diameter of the outer ring 2 is 110 mm and the inner diameter of the inner ring 3 is 80 mm (specifically, in the case of a ball bearing with a bearing number of 6916, in this case the pitch circle diameter dm = 95.5 (mm) and the diameter of the balls 5 Da = 1 1 / 32 inch = 8.73125 (mm)), the number Z of balls 5 is 19. When applying the present invention to a ball bearing 1 having the same dimensions, the number Z of balls 5 can be 13 or less (r = 2.6 or more), preferably 8 or less (r = 4.3 or more).

[0044] Furthermore, in the case of a standard ball bearing 1 in which the outer diameter of the outer ring 2 is 62 mm and the inner diameter of the inner ring 3 is 40 mm (specifically, in the case of a ball bearing with a designation number of 6908, in this case the pitch circle diameter dm = 51 (mm) and the diameter of the balls 5 Da = 1 / 4 inch = 6.35 (mm)), the number Z of balls 5 is 14. When applying the present invention to a ball bearing 1 having the same dimensions, the number Z of balls 5 can be 9 or less (r = 2.8 or more), and preferably 6 or less (r = 3.2 or more).

[0045] As described above, the present invention is applicable to ball bearings having bearing series codes of 67, 68, 69, 60, 62, 63, and 64 as defined in Japanese Industrial Standards JIS B1513-1995, "Nominal Designations of Rolling Bearings," for example, and having an inner diameter of the inner ring 3 of 30 mm to 90 mm.

[0046] As shown in Figure 3, a pocket 15 for accommodating a ball 5 is formed between each pair of cage claw portions 12. The inner surface of the pocket 15 (the surfaces of the cage claw portions 12 and the cage annular portion 11 facing the surface of the ball 5) is a concave spherical surface that conforms to the surface of the ball 5. An oil groove 16 is formed on the inner surface of the pocket 15, extending radially (in the direction perpendicular to the plane of the drawing) at an axial position (in the vertical direction in the drawing) corresponding to the center O of the ball 5.

[0047] As shown in Figures 1 and 4, the portion of the retainer annular portion 11 that connects the retainer claw portions 12 between circumferentially adjacent pockets 15, specifically, the portion of the retainer annular portion 11 that connects the retainer claw portion 12 on the other circumferential side (right side in the upper part of the figure) of the pair of retainer claw portions 12 that form the pocket 15 on one circumferential side (left side in the upper part of the figure) of the pockets 15 that are adjacent in the circumferential direction (left-right direction in the upper part of the figure), and the retainer claw portion 12 on the one circumferential side (left side in the upper part of the figure) of the pair of retainer claw portions 12 that form the pocket 15 on the other circumferential side (right side in the upper part of the figure) (hereinafter referred to as the "inter-pocket portion 17"), is formed in a flat plate shape that extends circumferentially with a constant axial thickness.

[0048] 2, a side surface 18 on the other axial side (left side in the figure) of the inter-pocket region 17 of the cage annular portion 11 is formed as a flat surface perpendicular to the axial direction. Here, the side surface 18 on the other axial side (left side in the figure) of the inter-pocket region 17 of the cage annular portion 11 is formed at a position corresponding to the axial position of the end of one axial side (right side in the figure) of the inner ring raceway surface 9. Specifically, the inter-pocket region 17 of the cage annular portion 11 is formed so that the axial deviation between the axial position of the side surface 18 on the other axial side (left side in the figure) of the inter-pocket region 17 of the cage annular portion 11 and the axial position of the end of one axial side (right side in the figure) of the inner ring raceway surface 9 is within 5% of the diameter of the balls 5.

[0049] As shown in Figure 3, the axial thickness a (mm) of the inter-pocket region 17 of the retainer annular portion 11 is set to a size that satisfies a > b and a < c, where b (mm) is the axial thickness of the portion of the retainer annular portion 11 corresponding to the bottom of the pocket 15, and c (mm) is the axial distance from the side surface 19 on one axial side (the lower side in the figure) of the retainer annular portion 11 to the center O of the ball 5.

[0050] In addition, the axial thickness a (mm) of the pocket-to-pocket region 17 of the retainer annular portion 11 is set to a size of 1 / 70 to 1 / 30 of the pitch circle diameter dm (mm) of the balls 5, and the axial thickness b (mm) of the portion of the retainer annular portion 11 corresponding to the bottom of the pocket 15 is set to a size of 1 / 62 to 1 / 26 of the pitch circle diameter dm (mm) of the balls 5.

[0051] 5 shows a drive device (e-Axle) for driving an electric vehicle that uses the above-described ball bearing 1. This drive device is a unit that integrates an electric motor 20 for driving an electric vehicle such as an EV (battery electric vehicle) or HEV (hybrid electric vehicle) with a reducer 21 that reduces the rotation speed of the electric motor 20.

[0052] This drive device has an electric motor 20 having a rotor shaft 22, an input shaft 23 of a reducer 21 to which the rotation of the rotor shaft 22 is input, an output shaft 24 of the reducer 21, an input gear 25 fixed to the input shaft 23, an output gear 26 fixed to the output shaft 24, an intermediate gear 27 that reduces the rotation of the input gear 25 and transmits it to the output gear 26, and an intermediate shaft 28 fixed to the intermediate gear 27 so as to rotate integrally with the intermediate gear 27.

[0053] The rotor shaft 22 is a drive shaft that is rotated and driven by the electric motor 20. The output shaft 24 of the reducer 21 is connected to wheels (not shown) of the electric vehicle. The input shaft 23, intermediate shaft 28, and output shaft 24 are arranged in parallel with a gap between them. The input gear 25, intermediate gear 27, and output gear 26 are each helical gears. This drive device reduces the rotation of the electric motor 20 by transmitting it to the input gear 25, intermediate gear 27, and output gear 26 in that order, and then outputs the reduced rotation from the output shaft 24 to the wheels (not shown).

[0054] The rotor shaft 22, input shaft 23, and intermediate shaft 28 are rotatably supported by the ball bearing 1 of this embodiment. On the other hand, the output shaft 24 is rotatably supported by a tapered roller bearing 29. The ball bearing 1 is lubricated by a continuous supply of lubricating oil from outside the bearing while the bearing is rotating. The lubricating oil can be supplied to the ball bearing 1 by splash lubrication, in which the rotation of the output gear 26 splashes the lubricating oil up and splashes the lubricating oil onto the ball bearing 1; jet lubrication, in which lubricating oil pumped by an oil pump (not shown) is sprayed onto the ball bearing 1 from a nozzle; or air-oil lubrication, in which lubricating oil mixed with compressed air is sprayed onto the ball bearing 1 from a nozzle.

[0055] By the way, the ball bearing 1 shown in FIG. 2 is calculated by multiplying the dmn value (pitch circle diameter dm (mm) of the ball 5 by the rotation speed n (min) －１When used in a high-speed rotation range where the axial load (R) is 650,000 or more (or 800,000 or more in some cases), the balls 5 and cage 6 rotate at high speed, causing the air in the bearing space 4 between the outer ring 2 and the inner ring 3 to move circumferentially at high speed, creating an air curtain. This air curtain makes it difficult for lubricating oil supplied from outside the bearing to enter the bearing space 4, which can lead to the problem of starvation (running out of lubricating oil inside the bearing).

[0056] Here, in the case of a ball bearing 1 that supports the main spindle of a machine tool or the like, it is possible to prevent starvation by increasing the amount of lubricating oil supplied to the ball bearing 1. However, in the case of a ball bearing 1 used in a drive unit (e-Axle) for running an electric vehicle as shown in FIG. 5, it is difficult to increase the amount of lubricating oil supplied to the ball bearing 1.

[0057] 2 rotates at high speed, the lubricating oil supplied to the bearing space 4 from outside the bearing tends to move radially outward due to the action of centrifugal force, while the contact area between the balls 5 and the inner ring raceway surface 9 is located at the radially innermost part of the bearing space 4, and therefore the contact area between the balls 5 and the inner ring raceway surface 9 is less susceptible to lubrication from outside the bearing than other parts. Therefore, when the ball bearing 1 is used at high speeds, the area where the lubricating oil is most likely to be insufficient is the contact area between the balls 5 and the inner ring raceway surface 9, and if this contact area between the balls 5 and the inner ring raceway surface 9 can be effectively lubricated, it will be possible to effectively prevent starvation at high speeds.

[0058] To address the above problem, the ball bearing 1 of this embodiment has a longer circumferential spacing between circumferentially adjacent balls 5, as shown in Figure 1, and a thinner axial thickness at inter-pocket region 17 of retainer annular portion 11, as shown in Figure 2, so that when viewed from the radial outside, only a small portion of inner ring raceway surface 9 is covered by retainer 6. This means that lubricating oil supplied from outside the bearing can easily reach the contact areas between balls 5 and inner ring raceway surface 9.

[0059] 1, the circumferential distance between adjacent balls 5 is long, so the time intervals at which the balls 5 shown in FIG. 2 pass the same position on the inner ring raceway surface 9 are long. Therefore, insufficient lubrication is unlikely to occur at the contact points between the balls 5 and the inner ring raceway surface 9.

[0060] That is, when the balls 5 shown in FIG. 2 roll on the inner ring raceway surface 9, the balls 5 push aside the lubricating oil present on the inner ring raceway surface 9. If the next ball 5 then rolls at the same position on the inner ring raceway surface 9 before the amount of lubricating oil present on the inner ring raceway surface 9 has been restored, there is a risk that the balls 5 and the inner ring raceway surface 9 will come into direct contact without any intervening lubricating oil. To address this problem, in this embodiment, as shown in FIG. 1 , the circumferential spacing between adjacent balls 5 is long, so the time interval between the balls 5 shown in FIG. 2 passing the same position on the inner ring raceway surface 9 is long. This makes it easier for the amount of lubricating oil present on the inner ring raceway surface 9 to be restored between the time when the balls 5 roll on the inner ring raceway surface 9 and push aside the lubricating oil present on the inner ring raceway surface 9, and the time when the next ball 5 rolls at the same position on the inner ring raceway surface 9. As a result, it is less likely that the balls 5 and the inner ring raceway surface 9 will come into direct contact without any intervening lubricating oil, and insufficient lubrication is less likely to occur at the contact points between the balls 5 and the inner ring raceway surface 9.

[0061] 1, the circumferential spacing between adjacent balls 5 is long, so the gap inside the bearing through which lubricating oil can pass in the axial direction is large (the gap through which lubricating oil supplied from outside the bearing can pass axially through bearing space 4 between outer ring 2 and inner ring 3, avoiding balls 5 and cage 6 shown in FIG. 2). Therefore, even when ball bearing 1 is used in the high-speed rotation range, a large amount of lubricating oil passes through bearing space 4, making it easy to ensure a sufficient amount of lubricating oil to lubricate the contact area between balls 5 and inner ring raceway surface 9.

[0062] Due to the synergistic effect of the above-mentioned actions, the ball bearing 1 of this embodiment can effectively lubricate the contact area between the balls 5 and the inner ring raceway surface 9 with lubricating oil supplied from outside the bearing, making it possible to effectively prevent starvation (depletion of lubricating oil inside the bearing) in the high-speed rotation range.

[0063] Furthermore, as shown in Fig. 1, the ball bearing 1 of this embodiment has a small number of balls 5, and therefore a small number of retainer claw portions 12 shown in Fig. 2. Therefore, when the ball bearing 1 is used in a high-speed rotation range, it is possible to suppress deformation of the retainer 6 due to centrifugal force such that each of the pairs of retainer claw portions 12 tilts radially outward.

[0064] 2 rotates at high speed, centrifugal force acts on each of the pairs of cage claws 12 extending axially from the cage annular portion 11, causing torsional deformation of the cage annular portion 11 due to the force transmitted from the cage claws 12. As a result, the entire cage 6 is deformed such that each of the pairs of cage claws 12 tilts radially outward. If the deformation of the cage 6 becomes large, the cage claws 12 may interfere with the balls 5, potentially causing abnormal heat generation. To address this problem, in this embodiment, as shown in FIG. 1, the number of balls 5 and the number of cage claws 12 are small, so that the total centrifugal force acting on the pairs of cage claws 12 when the cage 6 shown in FIG. 2 rotates at high speed is small, and the torsional deformation of the cage annular portion 11 caused by the force transmitted from the cage claws 12 is also small. Therefore, when the ball bearing 1 is used in the high-speed rotation range, it is possible to prevent the retainer 6 from being deformed by centrifugal force so that each of the pairs of retainer claw portions 12 tilts radially outward.

[0065] In addition, as shown in Figure 2, the ball bearing 1 of this embodiment has an oil gap 13 formed in a position close to the inner ring raceway surface 9 (between the inner circumference of the retainer annular portion 11 and the inner ring shoulder surface 10), making it possible to effectively lubricate the contact area between the balls 5 and the inner ring raceway surface 9.

[0066] In addition, in the ball bearing 1 of this embodiment, as shown in Figure 4, the side surface 18 between circumferentially adjacent pockets 15 of the retainer annular portion 11 is a flat surface perpendicular to the axial direction, so the air present in the space between circumferentially adjacent balls 5 shown in Figure 1 is less likely to be agitated by circumferential movement of the retainer annular portion 11. Therefore, the lubricating oil present in the bearing space 4 shown in Figure 2 can easily reach the contact area between the balls 5 and the inner ring raceway surface 9.

[0067] Furthermore, as shown in Figure 3, the ball bearing 1 of this embodiment allows lubricating oil to be introduced into the oil grooves 16 formed on the inner surface of the pocket 15, and this lubricating oil can effectively lubricate the contact areas between the cage 6 and the balls 5.

[0068] As shown in Fig. 5, the ball bearing 1 of this embodiment is suitable for use as a bearing supporting the rotor shaft 22 of an electric motor 20 used to drive an electric vehicle, or as a bearing supporting the input shaft 23 or intermediate shaft 28 of a reducer 21 that slows the rotation of that electric motor 20. That is, although the bearing supporting the rotor shaft 22 or the bearing supporting the input shaft 23 or intermediate shaft 28 of the reducer 21 rotates at an extremely high speed, the magnitude of the load supported by the bearing is very small (for example, a few percent or less) compared to the basic dynamic load rating of a standard ball bearing, and therefore it is possible to set the number of balls 5 to a small number, as shown in Fig. 1. For this reason, the use of the ball bearing 1 of this embodiment is particularly suitable.

[0069] In the above embodiment, a resin composition in which a fiber reinforcement material is added to a resin material is used as the material of the cage 6, but it is also possible to use mild steel (SPC) instead of the resin composition.

[0070] The embodiments disclosed herein should be considered to be illustrative in all respects and not restrictive. The scope of the present invention is defined by the claims, not by the above description, and is intended to include all modifications within the meaning and scope of the claims.

[0071] DESCRIPTION OF SYMBOLS 1 Ball bearing 2 Outer ring 3 Inner ring 4 Bearing space 5 Ball 6 Cage 7 Outer ring raceway surface 9 Inner ring raceway surface 10 Inner ring shoulder surface 11 Cage ring portion 12 Cage claw portion 13 Oil passage gap 15 Pocket 16 Oil passage groove 17 Area between pockets 18 Side surface 20 Electric motor 21 Reducer 22 Rotor shaft (rotating shaft) 23 Input shaft (rotating shaft) 28 Intermediate shaft (rotating shaft) a Axial thickness b Axial thickness c Axial distance O Center of ball

Claims

1. A ball bearing comprising: an outer ring (2) having an outer ring raceway surface (7) formed on its inner circumference; an inner ring (3) having an inner ring raceway surface (9) formed on its outer circumference; a plurality of balls (5) mounted at intervals in the circumferential direction in an annular bearing space (4) formed between the outer ring (2) and the inner ring (3) and in rolling contact with the outer ring raceway surface (7) and the inner ring raceway surface (9); and a retainer (6) for retaining the plurality of balls (5), wherein the bearing space (4) is in communication with the outside of the bearing so as to introduce lubricating oil from outside the bearing; the retainer (6) having a retainer annular portion (11) extending circumferentially on one axial side of the plurality of balls (5) and a plurality of pairs of retainer claw portions (12) extending from the retainer annular portion (11) to the other axial side; and a pocket (15) for accommodating each of the balls (5) formed between each pair of retainer claw portions (12), a ball bearing characterized in that an axial thickness a (mm) of a portion (17) connecting the retainer claw portions (12) between circumferentially adjacent pockets (15) of the retainer annular portion (11) satisfies a < c, where c (mm) is an axial distance from a side surface (19) on one axial side of the retainer annular portion (11) to the center (O) of the ball (5); and a ball-to-ball circumferential length ratio r, defined as r = (dm × π ÷ Z) / Da, is set to a value of 2.6 or more and 11.4 or less, where dm (mm) is the pitch circle diameter of the ball (5), Z (pieces) is the number of the balls (5), and Da (mm) is the diameter of the ball (5).

2. A ball bearing as described in claim 1, wherein the axial thickness a (mm) of the portion (17) connecting the retainer claw portions (12) between circumferentially adjacent pockets (15) of the retainer annular portion (11) satisfies a > b, where b (mm) is the axial thickness of the portion of the retainer annular portion (11) corresponding to the bottom of the pocket (15).

3. A ball bearing according to claim 1 or 2, wherein a cylindrical inner ring shoulder surface (10) is formed on the outer periphery of the inner ring (3), axially adjacent to the inner ring raceway surface (9) and radially facing the inner periphery of the retainer annular portion (11), and an annular oil passage gap (13) for allowing lubricating oil to pass is formed between the inner periphery of the retainer annular portion (11) and the inner ring shoulder surface (10).

4. A ball bearing as described in any one of claims 1 to 3, wherein the side surface (18) on the other axial side of the portion (17) connecting the retainer claw portions (12) between the circumferentially adjacent pockets (15) of the retainer annular portion (11) is a flat surface perpendicular to the axial direction.

5. A ball bearing as described in any one of claims 1 to 4, wherein an oil groove (16) extending radially at an axial position corresponding to the center (O) of the ball (5) is formed on the inner surface of the pocket (15).

6. A ball bearing according to any one of claims 1 to 5, wherein the axial thickness b (mm) of the portion of the retainer annular portion (11) corresponding to the bottom of the pocket (15) is set to a value between 1 / 62 and 1 / 26 of the pitch circle diameter dm (mm) of the ball (5).

7. A ball bearing as described in any one of claims 1 to 6, wherein the axial thickness a (mm) of the portion (17) connecting the retainer claw portions (12) between the circumferentially adjacent pockets (15) of the retainer annular portion (11) is set to a value between 1 / 70 and 1 / 30 of the pitch circle diameter dm (mm) of the balls (5).

8. A ball bearing according to any one of claims 1 to 7, wherein the cage (6) is formed from a resin composition in which a fiber reinforcing material is added to one of polyamide resin, polyether ether ketone resin, and polyphenylene sulfide resin.

9. A ball bearing according to any one of claims 1 to 7, wherein the cage (6) is made of mild steel.

10. A ball bearing according to any one of claims 1 to 9, which is used as a bearing supporting the rotating shaft (22) of an electric motor (20) for driving an electric vehicle, or as a bearing supporting the rotating shaft (23, 28) of a reducer (21) that reduces the rotation of the electric motor (20).

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