JP2024040109A - ball bearing - Google Patents

Abstract

To provide a ball bearing having excellent low torque without using a holder with a special shape.SOLUTION: A ball bearing comprises an inner ring 2 having an inner ring raceway groove on an outer periphery, an outer ring 3 having an outer ring raceway groove on an inner periphery, balls 4 interposed between the inner ring raceway groove and the outer ring raceway groove, and a holder 5 that holds the balls 4. The number Z of the balls 4 satisfies the condition of the following formula (1): Z≤((π×Dp)/2Da) ... (1), where Z represents the number of the balls 4 (Z is an integer larger than or equal to 3), Dp represents a pitch diameter of the ball 4 [unit: mm], and Da represents a diameter of the ball 4 [unit: mm].SELECTED DRAWING: Figure 1

Description

The present invention relates to a ball bearing, and particularly to a deep groove ball bearing for an e-Axle, which is an electric vehicle drive system mounted on an electric vehicle.

In recent years, from the perspective of environmental issues such as global warming, automobiles and industrial machinery are required to be even more energy efficient. Against this background, studies on improving the efficiency of motor systems are progressing in countries around the world. In order to develop a highly efficient motor system, it is essential to improve the efficiency of each component.

An example of the development of a high-efficiency motor system is the reduction in power consumption of electric vehicle drive systems as electric vehicles such as electric vehicles become popular. As an electric vehicle drive system, e-Axle is known, for example, which integrates the motor, inverter, gear box (reducer), etc. required for driving electric vehicles, which were previously arranged separately. . Here, electricity cost means the distance traveled per unit capacity of the electric energy source. In order to reduce the electricity consumption of an electric vehicle drive system, for example, it is necessary to reduce the torque of a rolling bearing such as a ball bearing, which is one of the components (reduce rotational torque).

Patent Document 1 discloses a ball bearing equipped with a retainer in which the distance between balls held in adjacent pockets is set to 2 times or more and 5 times or less than the ball diameter to improve ball insertion performance for the purpose of reducing torque. is listed. The cage of this ball bearing is a crown-shaped cage in which a plurality of pairs of claws forming the opening end of the pocket are formed in the circumferential direction. A height protrusion is formed. Patent Document 1 discloses that when rotating at high speed, the flat surface of the convex portion is continuous in the circumferential direction at the same height as the tip of the claw piece, so that no special stirring resistance is generated only in the claw piece. It is stated that the flow of lubricating oil can be adjusted and torque loss can be significantly reduced.

Patent Document 2 discloses a holding system that reduces the contact area and reduces the shear torque of lubricating oil by forming a recess in the pocket surface of the cage to provide a non-contact part with the balls for the purpose of reducing torque. The equipment is listed.

Patent No. 5348271 Patent No. 5602345

In Patent Document 1, for the purpose of reducing torque, the distance between the balls held in adjacent pockets is set to 2 times or more and 5 times or less of the ball diameter, and the cage has improved ease of ball insertion. Since the thickness of the convex portion between the pair of claw pieces becomes thicker and the weight of the retainer increases, deformation due to centrifugal force is likely to increase. Furthermore, in recent years, T/M (transmissions) and reduction gears for EVs have become mainstream in lean lubrication environments for the purpose of preventing environmental pollution and reducing torque. In a dilute lubrication environment, the retainer claws are not immersed in lubricating oil, so agitation by the retainer claws is unlikely to occur.

Patent Document 2 improves the pocket shape of an iron corrugated retainer, but due to the change in the pocket shape, stable holding of the balls decreases during high-speed rotation, etc., and deformation may occur.

Attempting to reduce the torque by changing the shape of the cage in this way tends to cause various problems, and it is desirable to easily reduce the torque without using a cage with a special shape.

The present invention was made in order to cope with such problems, and an object of the present invention is to provide a ball bearing with excellent low torque performance.

The ball bearing of the present invention includes an inner ring having an inner ring raceway groove on the outer periphery, an outer ring having an outer ring raceway groove on the inner periphery, balls interposed between the inner ring raceway groove and the outer ring raceway groove, and holding the balls. The ball bearing is characterized in that the number Z of the balls satisfies the condition of the following formula (1). Here, Dp and Da are both numerical values in "mm" units.
Z≦((π×Dp)/2Da)...(1)
Z: Number of balls mentioned above (Z is an integer of 3 or more)
Dp: pitch diameter of the above ball [unit: mm]
Da: Diameter of the above ball [unit: mm]

It is characterized in that the number Z of the balls satisfies the condition of formula (2) below.
4≦Z≦((π×Dp)/2Da)...(2)

When the groove curvature diameter of the inner ring raceway groove is Xi [unit: mm] and the groove curvature diameter of the outer ring raceway groove is Xo [unit: mm], the following conditions of formula (3) and formula (4) are satisfied. shall be.
0.515(π×Dp)/Z<Xi<0.54(π×Dp)/Z...(3)
0.535(π×Dp)/Z<Xo<0.55(π×Dp)/Z...(4)

The groove curvature diameter Xo of the outer ring raceway groove is 1.02 to 1.07 times the groove curvature diameter Xi of the inner ring raceway groove.

The ball bearing is characterized in that it includes seal members provided at openings at both axial ends of the inner ring and outer ring, and a lubricant supplied or sealed in the bearing internal space.

The seal member has a seal lip at an end, the inner ring or the outer ring has a seal sliding surface that slides in the circumferential direction with respect to the seal lip, and the seal lip is arranged in the bearing inner space. and a protrusion that creates a passage for the lubricant between the seal sliding surface and the seal lip that communicates with the outside, and the protrusion maintains a state of fluid lubrication between the seal lip and the seal sliding surface. It is characterized by being formed so as to.

A bearing for a drive system (e-Axle) that integrates a motor, an inverter that controls the motor, and a reducer that receives the rotation of the motor and transmits it to the drive shaft. .

The present invention is characterized in that the bearing rotatably supports at least one of the rotating shaft of the motor and the input shaft of the speed reducer.

In the ball bearing of the present invention, the number Z of balls satisfies the condition of formula (1) above, and the number of balls is reduced compared to the number N of balls in a conventional ball bearing, so the contact area between the balls and the bearing ring is In addition, when a lubricant is included, the number of times the lubricant is sheared is also reduced, leading to a reduction in rotational torque (torque loss) during rotation. As a result, the ball bearing of the present invention has excellent low torque performance without using a special cage.

Since the number of balls Z satisfies the condition of formula (2) above, the load can be supported by more than the predetermined number of balls, and although the rated load is reduced, the seal can prevent foreign matter from entering, resulting in low torque performance. It also has excellent bearing life.

In order to reduce torque, if the number of balls Z is smaller than before (for example, Z = (conventional number N) - 1), the load per rolling element will increase due to the decrease in the number of rolling elements. At low rotation speeds, torque loss due to slipping tends to be large. On the other hand, in the ball bearing, the groove curvature diameter Xi of the inner ring raceway groove satisfies the above formula (3), and the groove curvature diameter Xo of the outer ring raceway groove satisfies the above formula (4). The ratio of the groove curvature diameter of the inner and outer rings to the diameter of the balls used for this purpose (inner ring groove curvature diameter: 1.02 times, outer ring groove curvature diameter: 1.06 times) reduces torque loss due to slipping. be able to.

Further, when the difference between the inner ring curvature and the outer ring curvature becomes large, a difference in surface pressure between the inner and outer rings occurs, and there is a risk that the raceway ring on the side with higher surface pressure will peel off early. As a bearing, performance deteriorates significantly when flaking occurs on either the inner or outer ring, so the groove curvature diameter Xo of the outer ring raceway groove is 1.02 to 1.07 times the groove curvature diameter Xi of the inner ring raceway groove. Therefore, the difference in surface pressure between the inner and outer rings can be suppressed.

Since it has a sealing member provided at both axial end openings of the inner ring and outer ring, and a lubricant supplied or sealed into the bearing inner space, it is possible to prevent foreign matter from entering the bearing inner space, and the lubricant is equivalent to clean oil. It can be used in the following conditions. This suppresses peeling, indentation, and rolling fatigue that occur when foreign objects are caught in the bearing, and allows the bearing to have a life comparable to that of conventional bearings (when the number of balls is large).

The seal lip has a protrusion that creates a lubricant passage between the seal sliding surface and the seal lip that communicates between the internal space and the outside of the bearing, and the protrusion provides fluid lubrication between the seal lip and the seal sliding surface. Since the seal lip is formed so as to be in the bearing state, the friction coefficient of the sliding motion between the seal lip and the seal sliding surface can be significantly reduced, and the intrusion of foreign matter into the bearing inner space can be further suppressed. This results in particularly excellent low torque performance and long bearing life.

The ball bearing of the present invention is a bearing for a drive system (e-Axle) that integrates a motor, an inverter that controls the motor, and a reducer that receives the rotation of the motor and transmits it to the drive shaft. In particular, since the bearing rotatably supports at least one of the rotating shaft of the motor and the input shaft of the speed reducer, a relatively large load is not applied to the bearing. This makes it possible to reduce torque without changing bearing dimensions, and is expected to extend the cruising range of electric vehicles by reducing electricity consumption.

FIG. 1 is a plan view showing a first embodiment of the ball bearing of the present invention. FIG. 2 is a developed view of the periphery of adjacent pockets in the retainer of FIG. 1; FIG. 1 is a partial cross-sectional view of a ball bearing of the present invention. It is a partial sectional view showing a second embodiment of the ball bearing of the present invention. 5 is an enlarged view of the vicinity of the seal lip in FIG. 4. FIG. FIG. 3 is a partial front view of the seal lip viewed from the axial direction. FIG. 1 is a schematic diagram of an electric vehicle drive system.

In ball bearings, one way to reduce the torque without using a special cage is to reduce the number of balls. However, reducing the number of balls leads to a decrease in bearing rigidity, resulting in a decrease in the rated load. On the other hand, the present inventor has proposed that, for example, in a bearing used for a specific application (for example, an e-Axle bearing), by making the number Z of balls satisfy a predetermined condition, the negative effects caused by a decrease in the rated load can be avoided. We have discovered that it is possible to achieve low torque while avoiding this problem.

The ball bearing of the present invention is particularly suitable for applications requiring low torque without being subjected to high loads. For example, it can be used as a bearing for an electric vehicle drive system, and more specifically, it can be used as a bearing for an e-Axle. Note that details of the e-Axle will be explained with reference to FIG.

(First embodiment)
A first embodiment of the ball bearing of the present invention will be described based on FIG. 1. FIG. 1 is a plan view of a deep groove ball bearing as a ball bearing viewed from the axial direction. As shown in FIG. 1, a deep groove ball bearing (ball bearing) 1 includes an inner ring 2, an outer ring 3, balls 4 interposed between the inner ring 2 and the outer ring 3, and a crown-shaped retainer 5 that holds the balls 4. It is equipped with In FIG. 1, the cage 5 holds five balls 4 in five pockets so as to be able to roll freely at regular intervals in the circumferential direction. Note that the number of balls 4 in FIG. 1 is set to satisfy the following formula (1). Moreover, the deep groove ball bearing 1 does not have a seal member for suppressing foreign matter from entering the bearing inner space, and both ends of the inner ring 2 and the outer ring 3 in the axial direction are open.

As shown in FIG. 2, the retainer 5 has a plurality of opposing retaining claws 5b formed at a constant pitch in the circumferential direction on the circumferential upper surface of an annular resin main body 5a, and each of the opposing retaining claws 5b. The claws 5b are bent in a direction toward each other, and a pocket portion 5c for holding the ball is formed between the holding claws 5b. A flat portion 5d, which serves as a reference surface for rising up the holding claws 5b, is formed between the back surfaces of mutually adjacent holding claws 5b formed on the edges of the adjacent pocket portions 5c. As for the shape of the retainer 5, a general shape can be adopted except for the relationship with the number of pocket portions 5c. For example, as in Patent Document 1, the height of the holding claw and the height of the flat portion are not the same height.

Here, the number of balls in the ball bearing is set as follows.
(1) Setting of the bearing inner diameter, bearing outer diameter, and bearing width of the ball bearing (based on the JIS B 1512 dimensional standard), (2) Setting of the ball diameter Da, (3) Setting of the number of balls Z

For example, the P. of deep groove ball bearings used as e-Axle bearings. C. Assuming that D (the so-called pitch diameter Dp) is designed to be 45 mm to 65 mm, the bearing width B is 5 mm to 25 mm, and the ball diameter Da is 3 mm to 150 mm, the number of balls Z is usually 7 to 9. .

In the ball bearing of the present invention, the number Z of balls satisfies the condition of formula (1) below.
Z≦((π×Dp)/2Da)...(1)
Z: Number of balls (Z is an integer of 3 or more)

That is, the number of balls in the ball bearing of the present invention is one or more fewer than the number of balls in the conventional ball bearing. Thereby, the number of components of the ball bearing can be reduced, and rotational torque can be reduced while ensuring a certain degree of bearing rigidity. Additionally, the weight of the bearings is reduced, leading to lower electricity costs and lower fuel consumption. Furthermore, by reducing the stress generated when the balls fit between the raceway rings, defects in the roundness of the inner and outer rings are less likely to occur, and since the load per ball increases, the load on the bearing is reduced. Slippage can be suppressed even under low loads. In addition, by reducing the number of times the ball passes through the ball, it is expected that the creep phenomenon (inchworm phenomenon) will be reduced. Furthermore, even if the ball bearing of the present invention cannot reduce the viscosity of the lubricant due to its relationship with other parts (high viscosity lubricant must be used), it is possible to reduce torque by setting the number of balls. can.

For example, when the ball bearing of the present invention has a pitch diameter Dp of 45 mm to 65 mm, a bearing width B of 5 mm to 25 mm, and a ball diameter Da of 0.6 B, the number of balls Z is preferably 5 to 8. The number is more preferably 5 to 7, and even more preferably 5 to 6.

Moreover, in the ball bearing of the present invention, it is preferable that the number Z of balls satisfies the condition of the following formula (2).
4≦Z≦((π×Dp)/2Da)...(2)

For example, when the ball bearing of the present invention has pitch diameter Dp = 60 mm, bearing width B = 18 mm, and ball diameter Da = 0.6 B, the number Z of balls satisfying the above formula (2) is 5 or 6. , 7, or 8. For example, the maximum number of these (eight in this case) can be adopted. Further, when the pitch diameter Dp=45 mm, the bearing width B=25 mm, and the ball diameter Da=0.6B, the number Z of balls satisfying the above formula (2) is five.

By adopting the maximum number as the number Z while satisfying the above formula (1) or formula (2), the decrease in the rated load is suppressed to the minimum, low torque can be achieved, and the bearing life is also excellent.

Here, when the number of balls is reduced, the load per rolling element increases due to the decrease in the number of rolling elements, and in the case of low rotation, torque loss due to slipping increases. In order to reduce torque loss due to this slippage, it is preferable to set the ratio of the groove curvature diameter of the inner ring and outer ring raceway grooves to the diameter of the balls.

In the ball bearing 1' shown in FIG. 3, an inner ring raceway groove 2a having an arcuate cross section is formed along the circumferential direction on the outer periphery of the inner ring 2, and an outer ring raceway groove 3a having an arcuate cross section is formed on the inner periphery of the outer ring 3. It is formed along the direction. It is preferable that the groove curvature diameter Xi of the inner ring raceway groove 2a satisfies the condition of the following formula (3).
0.515(π×Dp)/Z<Xi<0.54(π×Dp)/Z...(3)
Moreover, it is preferable that the groove curvature diameter Xo of the outer ring raceway groove 3a satisfies the condition of the following formula (4).
0.535(π×Dp)/Z<Xo<0.55(π×Dp)/Z...(4)
Note that the inner ring raceway groove 2a and the outer ring raceway groove 3a each have a single curvature. A ball bearing 1' shown in FIG. 3 is a bearing that has a seal member 6 that seals the inner space of the bearing and is lubricated with a lubricant 7 such as lubricating oil or grease.

Regarding each of the groove curvature diameters Xi and Xo of the inner ring and outer ring raceway grooves, the larger the numerical value, the smaller the contact area between the raceway ring and the balls, which is advantageous for torque reduction. On the other hand, since the surface pressure against the load increases and rattling tends to increase, it is preferable to satisfy the conditions of the above formulas (3) and (4) especially in high-speed rotation applications. For example, when formulas (3) and (4) are satisfied, loss torque due to slipping can be reduced by about 10% while suppressing rattling (based on experimental results). Note that the balls are not limited to steel balls, and the same effect can be obtained even when ceramic balls are used.

Furthermore, it is preferable that the groove curvature diameter Xi of the inner ring raceway groove satisfies the condition of the following formula (5), and it is preferable that the groove curvature diameter Xo of the outer ring raceway groove satisfies the condition of the following formula (6). As a form that satisfies the following formulas (5) and (6), for example, the groove curvature diameter Xi of the inner ring raceway groove is 1.07 times the ball diameter Da, and the groove curvature diameter Xo of the outer ring raceway groove is the ball diameter. An example is 1.10 times Da.
0.53(π×Dp)/Z<Xi<0.54(π×Dp)/Z...(5)
0.54(π×Dp)/Z<Xo<0.55(π×Dp)/Z...(6)

If the difference between the inner ring curvature and the outer ring curvature becomes large, a difference in surface pressure will occur between the inner and outer rings, and there is a risk that the raceway ring on the side with higher surface pressure will peel off early. Considering this, it is preferable that the groove curvature diameter Xo of the outer ring raceway groove is 1.02 to 1.07 times the groove curvature diameter Xi of the inner ring raceway groove. Thereby, the surface pressure difference can be reduced to 15% or less, for example. More preferably, the groove curvature diameter Xo of the outer ring raceway groove is 1.02 to 1.03 times the groove curvature diameter Xi of the inner ring raceway groove.

In the deep groove ball bearing 1 shown in FIGS. 1 and 3, the shape of the cage is not limited to a crown shape, but may be a wave shape as described in the second embodiment.

(Second embodiment)
A second embodiment of the ball bearing of the present invention will be described based on FIG. 4. Note that the same configurations as in the first embodiment are denoted by the same reference numerals, and detailed description thereof will be omitted. FIG. 4 is an axial cross-sectional view of a deep groove ball bearing having a seal member as a ball bearing according to a second embodiment. As shown in FIG. 4, the deep groove ball bearing 11 has a pair of seal members 6', 6' at openings at both axial ends of the inner ring 2 and outer ring 3. The seal member 6' has one end fixed to the outer ring 3. Note that the seal member 6' may have one end fixed to the inner ring 2. The deep groove ball bearing 11 has a lubricant 7 in the bearing inner space. As the lubricant 7, lubricating oil or grease can be used. The lubricant 7 is supplied or sealed within the bearing space, and lubricates the raceway surface and the like.

Here, if the number of balls is reduced by one, for example, from the number Z of balls in a conventional ball bearing, there is a concern that the life of the bearing will be reduced. In contrast, in the deep groove ball bearing 11 of the second embodiment, both ends of the inner ring 2 and outer ring 3 in the axial direction are closed by the seal members 6', 6', thereby suppressing the intrusion of foreign substances into the bearing internal space. be done. As a result, rather than reducing the bearing life due to a reduction in the number of balls 4, the effect of suppressing peeling, indentation, and rolling fatigue that occurs when foreign objects are caught in the ball 4 is reduced compared to the conventional number of balls. It is possible to secure a bearing life equivalent to that of In addition, the number of times that shearing force is applied to the lubricant per bearing rotation is reduced compared to conventional bearings, which makes it difficult for the lubricant to heat up, contributing to a longer life of the lubricant and preventing lubricant leakage. Hateful.

A seal member in the ball bearing of the second embodiment will be described based on FIG. 5. FIG. 5 is an enlarged view of the vicinity of the seal lip in FIG. 4. As shown in FIG. 5, the sealing member 6' has a sealing lip 61 at the end. The inner ring 2 has a seal sliding surface 2b that slides in the circumferential direction with respect to the seal lip 61. Note that when the seal member 6' has one end fixed to the inner ring 2, the outer ring may have a seal sliding surface that slides in the circumferential direction with respect to the seal lip 61.

The seal lip 61 has, for example, a protrusion 62 that creates a lubricant passage (hereinafter also referred to as an "oil passage") between the seal sliding surface 2b and the seal lip 61 that communicates between the bearing internal space and the outside. are doing. Preferably, the protrusion 62 is formed so as to maintain fluid lubrication between the seal lip 61 and the seal sliding surface 2b. Further, it is preferable that the protrusion 62 has a shape and height that prevents foreign matter from passing through that may affect the life of the bearing. Thereby, it is possible to have sealing performance equivalent to that of a conventional contact seal. Note that the seal lip 61 is not limited to the protrusion 62, but may have a groove or a recess as long as it is formed so as to maintain fluid lubrication between the seal lip 61 and the seal sliding surface 2b.

With the above configuration, the lubricant in the passage is quickly dragged between the seal sliding surface and the seal lip due to the wedge effect as the bearing rotates, thereby promoting the formation of an oil film of the lubricant between the seal sliding surface and the seal lip. For this reason, a very thin oil film is formed between the seal lip and the seal sliding surface, and the bearing is completely separated by the lubricant and can be operated without direct contact (i.e. fluid lubrication), making it possible to operate the bearing without contact. The rotational torque can be reduced to almost the same as that of a seal.

Note that if a non-contact seal is used as the seal member, it is possible to reduce the oil agitation resistance, but with a general non-contact seal, the gap between the inner ring and the seal member is large relative to foreign objects, so foreign objects Invasion cannot be adequately suppressed. On the other hand, by employing the above-mentioned sealing member, an ideal environment is created that achieves both low torque and suppression of foreign matter intrusion.

Details of the seal member shown in FIG. 5 will be explained based on FIG. 6. FIG. 6 is a partial front view of the seal lip portion shown in FIG. 5 when viewed in the axial direction from the bearing inner space side in an independent and natural state. Here, the natural state refers to a state in which no external force is acting on the seal member in its independent state, that is, a state in which the seal member is not deformed by external force (hereinafter, this state is simply referred to as "natural state"). (referred to as “state”).

As shown in FIG. 6, the seal lip 61 has a tip 63 that defines the inner diameter of the seal lip 61 in its natural state. As shown in FIGS. 5 and 6, the protrusion 62 extends in a direction perpendicular to the circumferential direction. The protrusion 62 extends to the tip 63 of the seal lip 61 and is formed over the entire range with a radial interference between the protrusion 62 and the seal sliding surface 2b.

The protrusions 62 are arranged at regular intervals d in the circumferential direction. When the appearance of the seal lip 61 is viewed from the axial direction, a plurality of protrusions 62 appear radially arranged in the circumferential direction at a constant pitch angle θ corresponding to the interval d. Note that the radiation center is located on the central axis (coinciding with the bearing central axis) of the seal member 6' (not shown).

The distance d between the protrusions 62 adjacent to each other in the circumferential direction and the circumferential width W of the protrusions 62, together with the fact that the radially arranged protrusions 62 are located near the tip 63 of the seal lip 61, make the seal The lip 61 can come into sliding contact with the seal sliding surface 2b only on each projection 62, and an oil passage Co (see FIG. 5) is always created between each projection 62.

It is preferable that the height h of the protrusion 62 is set to 0.05 mm. In terms of design, this height h is the value at the highest position within the range where it can come into sliding contact with the seal sliding surface 2b. This position is also where the interference set between each protrusion 62 and the seal sliding surface 2b is maximum. Since the amount of deformation of the protrusion 62 during bearing operation can be ignored, the gap between the seal lip 61 and the seal sliding surface 2b in the direction perpendicular to the seal sliding surface 2b (including the oil passage Co) is The width at the narrowest point in the direction orthogonal to the surface 2b is equivalent to the height h of the protrusion 62, and does not substantially exceed 0.05 mm. For this reason, even if foreign matter with a particle size exceeding 50 μm is contained in the external lubricating oil, it is considered that the foreign matter will almost never pass through the oil passage Co. As a result, further extension of bearing life is expected.

It is preferable that the protrusion 62 has an R shape (approximately semi-cylindrical shape when viewed in the axial direction) that gradually approaches the seal sliding surface 2b from both ends of the circumferential width W toward the center of the circumferential width. This rounded shape is provided over the entire length of the protrusion 62 in the radial direction. Therefore, a region where the protrusion 62 and the seal sliding surface 2b can come into sliding contact exists in a linear manner on the virtual axial plane Pax passing through the center of the circumferential width of the protrusion 62. The center of curvature of the R shape of the protrusion 62 is on the virtual axial plane Pax. As a result, when the seal lip approaches the seal sliding surface in the presence of lubricant, an oil film is more likely to be instantaneously formed between the seal sliding surface of the inner ring and the seal lip due to the wedge effect.

In addition, in the deep groove ball bearing 11 of FIG. 4, the shape of the seal member 6' is not limited to the shape described above.

The structure of an e-Axle will be described as an example of an electric vehicle drive system using FIG. 7. The e-Axle is a drive unit that integrates a motor, an inverter, and a speed reducer. As shown in FIG. 7, the electric vehicle drive system 8 includes a motor 81 and a deceleration system that transmits the power of the motor 81 from a first shaft (input shaft) _S1 to a third shaft _S3 , which is a drive shaft, while reducing the speed. It is equipped with a machine 82. The inside of the housing 83 of the reduction gear 82 is filled with lubricating oil. The electric vehicle drive system 8 also includes an inverter that controls the motor 81 (not shown). Note that the first shaft _S1 may be a shaft of a separate member connected to the motor support shaft _Sm , or a single shaft may serve as both shafts.

The reducer 82 includes a first shaft S ₁ , a second shaft (counter shaft) S ₂ , a third shaft (output shaft) S ₃ , an input gear G ₀ , a first counter gear G ₁ , a second counter gear G ₂ , It has a drive gear _G3 . An input gear _G0 is provided on the first shaft _S1 . The second shaft _S2 is provided with a first counter gear _G1 and a second counter gear _G2 . A drive gear _G3 is provided on the third shaft _S3 .

The power of the motor 81 is directly transmitted to the first shaft S1, and the first shaft _S1 rotates. Input gear _G0 meshes with first counter gear _G1 . Further, the second counter gear _G2 meshes with the drive gear _G3 . The driving force of the first counter gear _G1 is decelerated (accelerated) by a combination of the second counter gear _G2 of the second shaft _S2 and the drive gear _G3 of the third shaft _S3 , and is transferred from the third shaft _S3. Output.

The electric vehicle drive system 8 further includes rolling bearings 12, 13, 14, 15, 16, 17, 18, and 19. The ball bearing of the present invention can be employed as at least one of the rolling bearings 12, 13, 14, 15, 16, and 17. The rolling bearings 18 and 19 are tapered roller bearings.

One end and the other end of the motor support shaft S _m are rotatably supported by rolling bearings 12 and 13, respectively. One end of the first shaft S ₁ on the motor 81 side is rotatably supported by a rolling bearing 14 , and the other end is rotatably supported by a rolling bearing 15 . The end of the second shaft _S2 on the side of the first counter gear _G1 is rotatably supported by a rolling bearing 16, and the end of the second shaft S2 on the side of the second counter gear _G2 is rotatably supported by a rolling bearing 17. has been done. Further, one end and the other end of the third shaft _S3 are rotatably supported by rolling bearings 18 and 19, respectively.

When the ball bearing of the present invention is used as a bearing for an e-Axle, it is preferably used as a bearing for a shaft other than the drive shaft (tapered shaft), which is the final shaft to which the largest load is applied. Since the decelerated second shaft _S2 is also subjected to a relatively large load, it is particularly preferable that the ball bearing of the present invention is applied to the motor support shaft _Sm and the first shaft _S1 . Note that in the case of a relatively low load, it can also be applied to the second axis _S2 .

Reduction gears are generally lubricated using an oil bath method in which the housing is filled with lubricating oil, and hard foreign matter such as gear wear debris tends to get mixed into the lubricating oil. When a bearing without a seal member is used as a bearing in a motor or a speed reducer, hard foreign matter easily enters the space inside the bearing. In that case, there is a risk that the steel balls will step on the hard foreign matter that has entered the bearing internal space, leaving impressions on the raceway surface and leading to early failures. Therefore, if a bearing without a seal member is used in the reduction gear of an e-Axle, the life of the bearing may be significantly reduced.

The ball bearing of the present invention has a seal member provided at openings at both axial ends of an inner ring and an outer ring, and a lubricant supplied or sealed in the bearing inner space, and can be used as an e-Axle bearing. Particularly preferred. This makes it difficult for hard foreign objects to get mixed into the lubricating oil, contributing to longer bearing life. In addition, since it is difficult to apply excessively large loads to the e-Axle, a decrease in bearing rigidity due to a reduction in the number of balls in the bearing compared to the conventional one is less likely to be a problem, making it possible to maximize the benefits of lower torque. can. Note that when the ball bearing of the present invention is used as a bearing for an e-axle, in which hard foreign matter such as gear wear powder is difficult to mix into lubricating oil, it can also be used in a form without a seal member.

Although deep groove ball bearings are shown as ball bearings in FIGS. 1 to 7, other ball bearings such as angular contact ball bearings can also be used.

Since the ball bearing of the present invention has excellent low torque properties, it can be suitably used in various applications requiring low fuel consumption and low electricity consumption. In particular, it can be suitably used as a bearing for an e-Axle, which requires low electricity consumption and requires a relatively small load on the bearing.

1, 1', 11 Ball bearing (deep groove ball bearing)
12, 13, 14, 15, 16, 17, 18, 19 Rolling bearing 2 Inner ring 2a Inner ring raceway groove 2b Seal sliding surface 3 Outer ring 3a Inner ring raceway groove 4 Ball 5 Cage 6, 6' Seal member 61 Seal lip 62 Projection 63 Tip 7 Lubricant Co Oil passage 8 Electric vehicle drive system 81 Motor 82 Reducer 83 Housing

Claims (8)

A ball comprising an inner ring having an inner ring raceway groove on the outer periphery, an outer ring having an outer ring raceway groove on the inner periphery, balls interposed between the inner ring raceway groove and the outer ring raceway groove, and a retainer that holds the balls. A bearing,
A ball bearing characterized in that the number Z of the balls satisfies the condition of the following formula (1).
Z≦((π×Dp)/2Da)...(1)
Z: Number of balls (Z is an integer of 3 or more)
Dp: pitch diameter of the ball [unit: mm]
Da: diameter of the ball [unit: mm] The ball bearing according to claim 1, wherein the number Z of the balls satisfies the following formula (2).
4≦Z≦((π×Dp)/2Da)...(2) 3. The ball bearing according to claim 1 or 2, characterized in that, when a groove curvature diameter Xi [unit: mm] of the inner ring raceway groove and a groove curvature diameter Xo [unit: mm] of the outer ring raceway groove are defined as, the conditions of the following formulas (3) and (4) are satisfied:
0.515(π×Dp)/Z<Xi<0.54(π×Dp)/Z (3)
0.535(π×Dp)/Z<Xo<0.55(π×Dp)/Z (4) The ball bearing according to claim 3, wherein the groove curvature diameter Xo of the outer ring raceway groove is 1.02 to 1.07 times the groove curvature diameter Xi of the inner ring raceway groove. The ball bearing according to claim 1 or 2, characterized in that the ball bearing includes a seal member provided at openings at both axial ends of the inner ring and the outer ring, and a lubricant supplied or sealed in the bearing inner space. ball bearings. The sealing member has a sealing lip at an end,
The inner ring or the outer ring has a seal sliding surface that slides in a circumferential direction with respect to the seal lip,
The seal lip has a protrusion that creates a passage for the lubricant between the seal sliding surface and the seal lip that communicates between the inner space of the bearing and the outside,
6. The ball bearing according to claim 5, wherein the protrusion is formed so as to maintain fluid lubrication between the seal lip and the seal sliding surface. Claim 1 or claim 1, wherein the bearing is for a drive system in which a motor, an inverter that controls the motor, and a reducer that receives rotation of the motor and transmits it to a drive shaft are integrated. Ball bearing according to item 2. 8. The ball bearing according to claim 7, wherein the bearing rotatably supports at least one of a rotating shaft of the motor and an input shaft of the speed reducer.

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