JP5862162B2 - Tandem angular contact ball bearings - Google Patents

Description

  The present invention relates to an improvement in a tandem angular ball bearing that is incorporated in a rotary machine such as a differential gear and a transfer device for an automobile and supports a rotating shaft that rotates in a state where a radial load and an axial load are applied. Specifically, it is intended to realize a tandem angular ball bearing that facilitates reduction in size and weight and reduction in rotational resistance (dynamic torque) while ensuring the required durability.

  Tandem angular ball bearings are used instead of tapered roller bearings to keep the rotational resistance (dynamic torque) of the rotating support that supports the pinion shaft, which forms the differential gear for automobiles, in the differential case. This is conventionally considered as described in Patent Documents 1 to 5. Tandem angular ball bearings do not involve large sliding contact as in the case of tapered roller bearings during operation, so that the dynamic torque can be kept low and the resistance of the differential gear can be reduced. And the improvement of the performance centering on the acceleration performance and the fuel consumption performance of the automobile equipped with this differential gear can be achieved.

  FIG. 3 shows an example of a conventional structure of a rotation support device for a pinion shaft for a differential gear, which is composed of only ball bearings (without using tapered roller bearings) described in Patent Document 1. . Since the structure and operation of the entire differential gear are well known and described in Patent Documents 1 to 4, illustration and detailed description are omitted, and only the structure of the rotation support device portion will be described below. To do. A pair of ball bearings 1 and 2 are arranged in the differential case in a state of being separated from each other in the axial direction, and the pinion shaft 3 is supported by these ball bearings 1 and 2. These ball bearings 1 and 2 are angular ball bearings in which the balls have contact angles, and the contact angles of these ball bearings 1 and 2 are opposite to each other. Therefore, the pinion shaft 3 is rotatably supported in the differential case in a state where not only a radial load but also an axial load in both directions is supported.

  Among the ball bearings 1 and 2, a tandem angular ball bearing, which is a subject of the present invention, is used as the ball bearing 1 disposed on the pinion gear 4 side (pinion gear side) to which a relatively large radial load and axial load are applied. doing. On the other hand, the ball bearing 2 on the side opposite to the pinion gear 4 (on the side opposite to the pinion gear) that supports only a relatively small radial load and axial load is of a single-row angular type. In addition, as described in Patent Documents 2 to 4, a structure in which not only the pinion gear side but also the anti-pinion gear side is a tandem angular ball bearing has been conventionally known. The ball bearing 1 on the pinion gear side supports an axial load in a direction away from a ring gear (not shown) meshed with the pinion gear 4 (rightward in FIG. 3) in addition to the radial load. On the other hand, the ball bearing 2 on the anti-pinion gear side supports an axial load in a direction approaching the ring gear (leftward in FIG. 3) in addition to the radial load.

  A pinion gear-side ball bearing 1 that is a tandem angular ball bearing includes an outer ring 5, an inner ring 6, a plurality of balls 7a and 7b, and a pair of cages 8a and 8b. Of these, the outer ring 5 is provided with double-row angular outer ring raceways 9a and 9b having different inner diameters on the inner peripheral surface. The inner diameters of these outer ring raceways 9a and 9b are larger in the outer ring raceway 9a on the pinion gear side and smaller in the outer ring raceway 9b on the anti-pinion gear side. Further, the inner ring 6 is disposed concentrically with the outer ring 5 on the inner diameter side of the outer ring 5, and the outer diameters of the outer peripheral surfaces of the inner ring 6 are different from each other at portions facing the outer ring raceways 9 a and 9 b. Double row angular type inner ring raceways 10a and 10b are provided. Regarding the outer diameters of these inner ring raceways 10a and 10b, the inner ring raceway 10a on the pinion gear side is larger, and the outer ring raceway 10b on the anti-pinion gear side is smaller.

  Further, each of the balls 7a, 7b has a plurality of balls in each row between the outer ring raceways 9a, 9b and the inner ring raceways 10a, 10b (in a parallel combination type). ) It is provided so as to be able to roll while being given a contact angle. The cages 8a and 8b have different diameters from each other, and hold the balls 7a and 7b in both rows in a freely rollable manner. The pitch circle diameters (PCD) of the balls 7a and 7b in both rows correspond to the differences in the diameters of the tracks 9a, 9b, 10a and 10b, and the row on the pinion gear side is larger than the row on the anti-pinion gear side. .

  When the pinion shaft 3 is actually rotatably supported by the tandem angular ball bearing 1 as described above, the outer ring 5 is fitted into the differential case by an interference fit. It is necessary to externally fit the inner ring 6 to the pinion shaft 3 by an interference fit. However, the inner fitting operation of the outer ring 5 and the outer fitting operation of the inner ring 6 cannot be performed with the tandem angular ball bearing 1 assembled in advance. The reason for this is to prevent the formation of indentations on the tracks 9a, 9b, 10a, 10b. For this reason, conventionally, the inner ring work of the outer ring and the outer ring work of the inner ring can be performed separately by the structure as described in Patent Document 4.

  4 to 6 show a second example of the conventional structure described in Patent Document 4. FIG. The outer ring 5a constituting the tandem angular ball bearing 1a is not provided with a groove shoulder on one axial side (the left side in FIGS. 4 and 6) of both the large diameter side and small diameter side outer ring raceways 9a and 9b (counter bore). The groove shoulder is provided only on the other side in the axial direction (the right side in FIGS. 4 and 6). On the other hand, the inner ring 6a is provided with groove shoulders on both sides in the axial direction of both the large diameter side and small diameter side inner ring raceways 10a and 10b. Further, both the large-diameter side and small-diameter side retainers 8a and 8b hold the balls 7a and 7b in the pockets 11a and 11b, respectively, and the balls 7a and 7b are moved from the pockets 11a and 11b. It is configured so that it can be prevented from coming out radially outward.

  When assembling the ball bearing 1a, first, an inner ring side assembly 12 as shown by a solid line in FIG. 5 is assembled. For this purpose, as indicated by the chain line in the figure, the balls 7a and 7b are held in the pockets 11a and 11b of the cages 8a and 8b, and then as indicated by arrows in FIG. The balls 7a and 7b held by the two cages 8a and 8b are elastically deformed on the outer diameter side of the inner ring 6a, and the inscribed circles of the balls 7a and 7b are formed. The inner ring 6a is made to enter from the other axial side (the right side in FIG. 5) while elastically expanding the diameter. Then, as indicated by solid lines in FIG. 5, the balls 7a and 7b held by the cages 8a and 8b are assembled to the outer diameter side of the inner ring raceways 10a and 10b.

  As shown in FIG. 6, the inner ring side assembly 12 assembled in this manner is inserted into the inner diameter side of the outer ring 5a, which has been fitted and fixed to a differential case in advance, and is tandem angular type. The assembly of the ball bearing 1a is completed. Of the two outer ring raceways 9a, 9b formed on the inner peripheral surface of the outer ring 6a, there is no groove shoulder on the one side portion in the axial direction on the entry side of the balls 7a, 7b. Such an insertion operation of the inner ring side assembly 12 to the inner diameter side of the outer ring 5a can be performed smoothly. Thus, in the case of the second example of the conventional structure, the ball bearing 1a can be handled by being divided into two elements, the outer ring 5a and the inner ring side assembly 12. During the assembly operation between the differential case and the pinion shaft 3, the rolling surfaces of the balls 7a and 7b and the contact portions between the outer ring raceways 9a and 9b and the inner ring raceways 10a and 10b. It is possible to prevent excessive surface pressure from acting on the surface. As a result, it is possible to prevent indentation from being formed on each of the tracks 9a, 9b, 10a, and 10b. During the assembling work as described above, the rolling surfaces of the balls 7a and 7b are prevented from being damaged at the sharp corners present on the inner peripheral surface of the outer ring 5a or the outer peripheral surface of the inner ring 6a. In order to do this, among these peripheral surfaces, there is no non-differentiable (pointed) corner portion where the rolling surface of each of the balls 7a, 7b may come into contact during the assembly operation, and Further, a smooth surface that has been polished is described in Patent Document 5 and is conventionally known.

  In any structure, unlike the tapered roller bearings, the tandem angular ball bearings 1 and 1a as described above do not involve large sliding contact during operation, so that the dynamic torque can be kept low and the resistance of the differential gear can be reduced. Can be lowered. Moreover, since the balls 7a and 7b arranged in a double row support the radial load and the axial load generated at the meshing portion of the pinion gear 4 and the ring gear, it is possible to sufficiently secure the load capacity regarding the loads in both directions. . However, there is still room for improvement in terms of reducing the size and weight and reducing the dynamic torque while ensuring the required durability. This point will be described below.

  In order for the tandem angular ball bearings 1 and 1a, which are a kind of double-row rolling bearings, to perform their original functions, the rolling surfaces of the balls 7a and 7b arranged in both rows are in contact with each other. In addition, both the outer ring raceways 9a and 9b and the both inner ring raceways 10a and 10b must be normal (smooth surfaces). When any one of the rows causes damage such as separation on any of the rolling surface, the outer ring raceway, and the inner ring raceway, and the function of the row is impaired (the rolling fatigue life is reached), the ball bearing 1 1a is considered damaged (rolling fatigue life reached), requiring repair and replacement.

  By the way, during operation of a rotary machine such as a differential gear, the radial bearing and the axial load in the same direction are continuously applied to the ball bearings 1 and 1a for the same time on the balls 7a and 7b in both rows. On the other hand, unless special measures are taken, the rolling fatigue life of both rows will not be the same due to the difference in pitch circle diameter between the balls 7a, 7b in both rows. If there is a large difference in the rolling fatigue life between the two rows, the row with a long rolling fatigue life becomes excessive quality, which is disadvantageous in terms of cost reduction, size reduction / weight reduction, and torque reduction. In other words, the fact that only the rolling fatigue life of only one row does not lead to an improvement in the durability of the ball bearings 1 and 1a as a whole is significantly longer than the rolling fatigue life of the other row. This means that the cost required for extending the fatigue life is unnecessarily increased, which is disadvantageous in terms of reducing the cost of the ball bearings 1 and 1a. Further, when the conditions other than the dimensions such as the material and the surface properties are the same, the rolling fatigue life becomes longer as the dimensions are larger. Therefore, unnecessarily extending the rolling fatigue life of the one row is disadvantageous in terms of reducing the size and weight of the ball bearings 1 and 1a. Further, the dynamic torque of the ball bearings 1 and 1a increases as the dimensions of the ball bearings 1 and 1a increase when the conditions other than the dimensions are the same, so the rolling fatigue life of the one row is unnecessarily prolonged. This is also disadvantageous in terms of reducing the torque of the ball bearings 1 and 1a.

In addition, regarding the tandem angular ball bearing, the specifications of both rows are not necessarily the same. For example, the diameters of the balls of both rows (hereinafter referred to as “ball diameter”; the same applies to the claims) are made different. As a structure, the invention described in Patent Document 6 is known. The structure of the invention described in this Patent Document 6 is a four-point contact type on the row with the smaller pitch circle diameter, and the ball diameter with the smaller pitch circle diameter is the ball diameter with the larger pitch circle diameter. It is smaller than the diameter. In addition, the tandem angular ball bearing can be reduced in size while ensuring the ability to support an axial load in the opposite direction that acts rarely.
However, even with the invention described in the cited document 6 as described above, the difference in rolling fatigue life between the two rows cannot be reduced (rather increased).

JP 2004-169890 A JP 2004-245231 A JP 2009-138895 A JP 2004-124996 A International Publication No. 2011/062257 Pamphlet Special table 2010-534303 gazette

  In view of the circumstances as described above, the present invention was invented to realize a tandem angular ball bearing that is easy to reduce size and weight and reduce dynamic torque while ensuring the required durability. is there.

The tandem angular ball bearing of the present invention includes an outer ring, an inner ring, and a plurality of balls in the same manner as a conventionally known tandem angular ball bearing.
Of these, the outer ring is provided with two rows of outer ring raceways having different inner diameters on the inner peripheral surface.
The inner ring is disposed concentrically with the outer ring on the inner diameter side of the outer ring, and two rows of inner ring raceways having different outer diameters are provided on the outer peripheral surface.
Further, each of the balls can freely roll between the outer ring raceways and the inner ring raceways in a state where a contact angle in the same direction is given between the rows, a plurality for each row. Provided. In general, the balls in both rows are held by independent cages.

In particular, in the tandem angular contact ball bearing of the present invention, the pitch circle diameter provided between the outer ring raceway and the inner ring raceway on the smaller side of the outer ring raceway and the inner ring raceway in both rows. The ball diameter on the first row side where the diameter is small is made larger than the ball diameter on the second row side where the pitch circle diameter is large, which is provided between the outer ring raceway and the inner ring raceway on the larger diameter side.
Moreover, the internal gap (radial gap and axial gap) related to the balls on the first row side is made smaller than the internal gap related to the balls on the second row side.
In this way, by providing a difference in the internal gap between the ball on the first row side and the ball on the second row side, the magnitude of the load applied to the ball on the first row side in use is reduced. , Larger than the load applied by the ball on the second row side.
For example, the ratio of the load applied by the balls on the first row side to the load applied by the balls on the second row does not cause a large difference in rolling fatigue life for both rows. Although it is determined by computer analysis or experiment, it is 1.5 to 2.5 times (most preferably 2 times) as in the present invention .

According to the present invention configured as described above, it is possible to realize a tandem angular ball bearing that facilitates reduction in size and weight and reduction in dynamic torque while ensuring the required durability. The reason will be described below.
First, the required durability can be ensured by increasing the ball diameter on the first row side where the pitch circle diameter is small and increasing the load applied by the ball on the first row side. The area of the contact ellipse existing at the rolling contact portion between the rolling surface of the ball on the first row side and the outer ring raceway and the inner ring raceway with a large ball diameter is the rolling of the ball on the second row side with a small ball diameter. It is wider than the area of the contact ellipse on the surface. On the other hand, the rolling fatigue life of each part changes according to the surface pressure of each of these contact ellipses, and the rolling fatigue life becomes shorter as the surface pressure increases. The surface pressure increases as the applied load increases and the contact ellipse area decreases. Therefore, as in the present invention, if the load applied by the balls in the first row is made larger than the load applied by the balls in the second row, the changes in the surface pressure due to the respective elements are opposite to each other. By canceling out, it is possible to prevent a large difference in the surface pressure of the contact ellipse relating to the balls in both rows, and it is possible to prevent a great difference in the rolling fatigue life between these rows. In any row, it is possible to easily ensure the required durability while preventing the durability from being unnecessarily long (only one of the rows is remarkable).

Moreover, since the durability of any row can be prevented from being lengthened (remarkably compared to the durability of another row), the size of any row portion can be prevented from becoming unnecessarily large. This is advantageous in terms of reducing the size and weight of the tandem angular contact ball bearing.
Further, since the size of any one of the row portions can be kept small and the size and weight can be reduced, this is advantageous in terms of reducing the torque of the tandem angular ball bearing.
In addition, when it is assumed that the integrated value of the surface pressure applied to the pitch circle diameter and the contact ellipse portion (the sum of the preload and the load) is the same, the ball diameter is large due to the influence of the spin loss and the like in the contact ellipse portion. As a result, the dynamic torque of the row of balls increases. Further, the dynamic torque increases in proportion to the pitch circle diameter. In the case of the present invention, since the pitch circle diameter of the first row of balls having a large ball diameter is small, an increase in dynamic torque caused by increasing the ball diameter of the first row can be kept low.

BRIEF DESCRIPTION OF THE DRAWINGS The fragmentary sectional view of the tandem angular ball bearing which shows one example of embodiment of this invention. The schematic sectional drawing which shows an example of the use condition of the tandem angular ball bearing of this invention. Sectional drawing of the rotation support part of the pinion shaft of a differential gear incorporating a tandem angular ball bearing. Sectional drawing of the tandem angular ball bearing which shows the 2nd example of conventional structure. Sectional drawing which shows the state which assembles the inner ring | wheel side assembly which comprises the tandem angular ball bearing of this 2nd example. Sectional drawing which shows the state which uses this inner ring | wheel side assembly and an outer ring | wheel as a tandem angular ball bearing.

An example of an embodiment of the present invention will be described with reference to FIG. The tandem angular ball bearing 1b of this example includes an outer ring 5b, an inner ring 6b, and a plurality of balls 7c and 7d.
Out of these, the outer ring 5b is provided with two rows of outer ring raceways 9c and 9d on the inner peripheral surface. In the case of this example, both the outer ring raceways 9c and 9d have not only different inner diameters but also different curvature radii r ₁ and r ₂ of the cross-sectional shapes. Specifically, the radius of curvature r ₁ of the outer ring raceway 9c on the side with the smaller inner diameter Rc is larger than the radius of curvature r ₂ of the outer ring raceway 9d on the side of the larger inner diameter Rd (r ₁ > r ₂ ). Further, portions of the inner peripheral surface of the outer ring 5b adjacent to the small diameter side of the outer ring raceways 9c, 9d are outer ring shoulder portions 13a, 13b, respectively. The height (radial dimension) of these outer ring shoulder portions 13a and 13b is such that the rolling surfaces of these balls 7c and 7d are independent of the axial load applied to the balls 7c and 7d when the ball bearing 1b is used. The size is large enough not to ride over the edges of the outer ring shoulder portions 13a and 13b.

  On the other hand, the portion of the inner peripheral surface of the outer ring 5b adjacent to the larger diameter side of the outer ring raceways 9c, 9d is gradually increased in the direction in which the inner diameter gradually increases as the distance from the outer ring raceways 9c, 9d increases. Inclined, partially conical concave counterbore portions 14a and 14b are provided. However, the inner diameters of the small-diameter end portions (the right end portion in FIG. 1) of both the counter bore portions 14a and 14b can be made slightly smaller than the maximum inner diameters of the two outer ring raceways 9c and 9d adjacent to each other. In this case, the slightly smaller extent is the radial internal clearance of the ball bearing 1b before the ball bearing 1b is assembled at a predetermined position (the outer ring 5b and the inner ring 6b can be displaced in the radial direction). It is preferable to make it smaller than 1/2 of (dimension). However, in consideration of manufacturing tolerances (for example, about 0.02 mm), the inner diameters of the end portions on the small diameter side of both the counterbore portions 14a and 14b are such that a part of the respective rolling surfaces corresponds to the inner ring raceways 10c and 10d. The diameter of the circumscribed circle of each of the balls 7c and 7d in the state of contact is not reduced. By restricting in this way, it is possible to easily assemble the inner ring side assembly 12 (see FIGS. 5 to 6) on the inner diameter side of the outer ring 5b (without causing damage such as scratches on each part). In this case, the state before the ball bearing 1b is assembled at a predetermined position is not the state of the ball bearing 1b alone, but the inner ring 5b constituting the ball bearing 1b is tightly fitted to a housing or the like. A state in which the inner ring 6b is externally fitted to the rotary shaft or the like by an interference fit. In short, the diameters of the outer ring 5b and the inner ring 6b are elastically changed by fitting with the mating member, but the preload is not yet applied to the balls 7c and 7d. In any case, a portion of the inner peripheral surface of the outer ring 5b where the rolling surfaces of the balls 7c and 7d may rub against each other during the assembly operation of the inner ring side assembly 12 is a non-differentiable angle. It has no part and is a polished smooth surface. For this purpose, it is preferable to finish the inner peripheral surface of the outer ring 5b with a grindstone.

The inner ring 6b is disposed concentrically with the outer ring 5b on the inner diameter side of the outer ring 5b, and two rows of inner ring raceways 10c and 10d are provided on the outer peripheral surface. These two inner ring raceways 10c, 10d are not different outer diameters from each other by, with regard radius of curvature r _3, r ₄ of the cross-sectional shape, are made different from each other. Specifically, the radius of curvature _{r 3} of the outer diameter Dc is smaller side of the inner ring raceway 10c, larger than the radius of curvature _{r 4} of the outer diameter Dd is larger side inner raceway 10d _(r 3> _{r 4)} to . Further, portions of the outer peripheral surface of the inner ring 6b adjacent to the large-diameter side of the inner ring raceways 10c and 10d are inner ring shoulder portions 15a and 15b, respectively. Regardless of the axial load applied to each of the balls 7c and 7d, the rolling surfaces of the balls 7c and 7d are not in contact with the inner ring shoulders 15a and 15b. It is large enough not to ride over the edge.

  Further, engaging portions 16a and 16b are provided on the outer peripheral surface of the inner ring 6b at portions adjacent to the small diameter side of the inner ring raceways 10c and 10d. The outer diameters of the engaging portions 16a and 16b are slightly larger than the minimum outer diameters of the inner ring raceways 10c and 10d adjacent to each other. The engaging portions 16a and 16b are combined with the inner ring 6b and the balls 7c and 7d through retainers 8a and 8b (see FIGS. 4 to 6 and omitted in FIG. 1) described later. It is necessary for preventing the inner ring 6b and the balls 7c and 7d from being separated in the state where the inner ring side assembly 12 is used. The engagement allowance (diameter in the radial direction) δ of both the engaging portions 16a and 16b, which is ½ of the difference between the outer diameter of the engaging portions 16a and 16b and the minimum inner diameter of the inner ring raceways 10c and 10d, The balls 7c and 7d can be combined with the inner ring 6b without damaging the rolling surfaces of the balls 7c and 7d, and the members 7c, 7d and 6b are not inadvertently separated after the combination. In view of the elasticity of both the cages 8a and 8b, the regulation is appropriately performed. Specifically, the engagement allowance δ of both the engaging portions 16a and 16b is restricted to about 2 to 15% of the diameter of each of the balls 7c and 7d that get over the engaging portions 16a and 16b. When the engagement allowance δ is less than 2%, separation of the members 7c, 7d, and 6b becomes uncertain, and when it exceeds 15%, it is difficult to combine these members 7c, 7d, and 6b. .

Further, the balls 7c, 7d have a plurality of contact angles in the same direction between the outer ring raceways 9c, 9d and the inner ring raceways 10c, 10d. In a state where θc and θd are given, they are provided to freely roll. Of the balls 7c and 7d, the ball diameter D of the ball 7c on the first row side is provided between the outer ring raceway 9c and the inner ring raceway 10c on the smaller diameter side, and the pitch circle diameter PCDc is small. _L a is provided between the outer ring raceway 9d and the inner ring raceway 10d of each larger diameter side, larger than the ball diameter D _S of the second row side of the ball 7d pitch diameter PCDd large (D _L> D _S ). Such balls 7c and 7d are arranged in the circumferential direction by cages 8a and 8b having shapes and dimensions that can prevent the balls 7c and 7d from coming out radially outward from the pockets 11a and 11b. It is kept at regular intervals.

The contact angles θc and θd of the balls 7c and 7d are appropriately regulated according to the load capacity related to the radial load and the axial load required for the ball bearing 1b. The values of the contact angles θc and θd in both rows may be the same or different, but in any case, they are regulated within a range of 10 to 40 degrees. Further, the a ball diameter _D L, _{D S} of the balls 7c, 7d, each track 9c, 9d, 10c, the radius of curvature about the cross-sectional shape of the _{10d r 1, r 2, r} 3, with respect to the _{r 4,} these Appropriate regulation is performed in consideration of the area of the contact ellipse existing at the rolling contact portion between each track 9c, 9d, 10c, 10d and the rolling surface of each ball 7c, 7d. The larger the area, the lower the surface pressure of the contact ellipse portion, which is advantageous from the viewpoint of ensuring the rolling fatigue life. Instead, the spin loss at the contact ellipse portion increases, and the ball bearing 1b. This is disadvantageous in terms of reducing dynamic torque.

In consideration of these facts, the outer ring raceways 9c, the radius of curvature _r 1 about the cross-sectional shape of 9d, _{r 2,} the respective balls 7c, ball diameter _D L of 7d, the range of 51 to 62% of _{D S} regulate. Further, the two inner ring raceways 10c, the radius of curvature _r 3 about the cross-sectional shape of the 10d, _{r 4,} the respective balls 7c, 7d of the ball diameter _D L, restricts the range of 50.5 to 58% of _{D S.} In any case, the curvature radius r ₁ (r ₂ ) related to the cross-sectional shape of the outer ring raceway 9c (9d) is calculated from the curvature radius r ₃ (r ₄ ) related to the cross-sectional shape related to the inner ring raceway 10c (10d). Also make it bigger. This is because the outer ring raceways 9c, 9d are concave in the circumferential direction, and the inner ring raceways 10c, 10d are convex in the circumferential direction. This is because the size of the contact ellipse relating to 7d is the same on both the outer ring raceways 9c, 9d and on both the inner ring raceways 10c, 10d. The dimensions D _L , D _S , r ₁ , r ₂ , r ₃ , and r ₄ are regulated so that the maximum value of the surface pressure of the rolling contact portion is suppressed to 4.2 GPa or less. In order to reduce the torque, as long as this condition is satisfied, it is preferable to increase the values of the respective radii of curvature r ₁ , r ₂ , r ₃ , r ₄ as much as possible.

Further, in the case of the ball bearing 1b of the present embodiment, the pitch circle diameter PCDc small, each ball diameter D _L is greater, the internal clearance for the first row side of the ball 7c, the pitch circle diameter PCDd large, each ball diameter D _S is small, it is smaller than the internal clearance for the second row side of the balls 7d. In the case of this example, the outer ring 5b and the inner ring 6b constituting the ball bearing 1b are brought closer to the direction in which the balls 7c and 7d are pressed, and the rolling surfaces of the balls 7c on the first row side are set. In the state where the outer ring raceway 9c and the inner ring raceway 10c are in light contact, there is a positive internal gap between the rolling surface of the balls 7d on the second row side and the outer ring raceway 9d or the inner ring raceway 10d. I try to do it. In the state where the ball bearing 1b is actually incorporated in the rotation support portion, the outer ring 5b and the inner ring 6b are pressed more strongly, and the internal clearance on the second row side is eliminated, and the balls 7c and 7d are preloaded. Is granted. In this state, the preload applied to each of the balls 7c and 7d increases on the first row side by the positive internal gap existing on the second row side in the intermediate state.

Ball bearing 1b of the present embodiment as described above, the pitch circle diameter PCDc is small, with a larger ball diameter D _L of the first row side of the ball 7c, increasing the preload of the first row side of the balls 7c, The magnitude of the load applied to the balls 7c on the first row side is increased. Therefore, the contact ellipse ball diameter D _L is greater, present in the rolling contact portion between the rolling surface of the first row side of the ball 7c, and each diameter is small outer ring raceway 9c and the inner ring raceway 10c is the ball diameter D _S is smaller than the contact ellipse related to the rolling surface of the ball 7d on the second row side (the area is wide). Among the ball bearings 1b, the rolling fatigue life of the surface which is in rolling contact with the mating surface with the relative rotation of the outer ring 5b and the inner ring 6b becomes shorter as the surface pressure of each contact ellipse portion becomes higher. In the case of the ball bearing 1b, it is possible to prevent a large difference in the rolling fatigue life between the first and second rows.

  In this way, it is possible to easily ensure the required durability while preventing the durability from becoming unnecessarily long (remarkably only in any column), and any column can be secured. The size of the part can be prevented from becoming unnecessarily large. In short, it is possible to prevent an increase in the size that does not lead to an improvement in the durability of the ball bearing 1b as a whole, but only to an improvement in the durability of only one of the rows, thereby reducing the size and weight of the tandem angular ball bearing. It is advantageous from the aspect of aiming.

  Further, since the size of any one of the row portions can be kept small and the size and weight can be reduced, this is advantageous in terms of reducing the torque of the tandem angular ball bearing. That is, useless enlargement leads to an increase in the resistance and inertial mass of the portions that are displaced relative to each other when the outer ring 5b and the inner ring 6b are relatively rotated. In the case of the ball bearing 1b of this example, as described above, Therefore, it is easy to reduce the torque of the ball bearing 1b.

Moreover, in the case of this example, since the ball diameter is the pitch circle diameter PCDc large first row of balls 7c is small, an increase in the dynamic torque caused by increasing the ball diameter D _L of the first row of balls 7c It can be kept low. That is, the rolling resistance of the balls constituting the ball bearing increases as the ball diameter increases. In other words, assuming that the pitch circle diameter and the integral value of the surface pressure applied to the contact ellipse part (the sum of the preload and the load) are the same, the spin loss at this contact ellipse part and the increase of the ball's inertial mass, etc. As a result, the dynamic torque of the row of balls increases as the ball diameter increases. Further, the dynamic torque increases in proportion to the pitch circle diameter. In the case of the present invention, the pitch circle diameter PCDc of the first row of balls 7c having a large diameter is made small, which is advantageous in terms of suppressing the dynamic torque of the ball bearing 1b as a whole.

The respective dimensions, that is, the inner diameters Rc and Rd of the outer raceways 9c and 9d, the outer diameter Dc and the outer diameter Dd of the inner raceways 10c and 10d, and the cross-sectional shapes of the raceways 9c, 9d, 10c, and 10d. curvature radius _{r 1, r 2, r 3} , r 4, wherein respective balls 7c, 7d of the contact angle .theta.c, [theta] d, the respective balls 7c, 7d ball diameter _D L, _{D S} of the such actions and above In order to obtain the effect in a high dimension, it is appropriately regulated within the above-mentioned range by computer analysis or experiment.

  Further, the number of the balls 7c and 7d in both the first and second rows is more preferable from the viewpoint of securing the load capacity of the ball bearing 1b. For example, a filling rate (D · Z / PCD × π), which is a value obtained by dividing the product of the ball diameter D and the number Z of balls by the circumference of the pitch circle (pitch circle diameter PCD × π), is 80 to 95. Regulate to the range of%. When the filling rate is less than 80%, it is difficult to secure the load capacity. On the other hand, when it exceeds 95%, it is difficult to assemble the cage (to secure the installation space for the pillars constituting the cage). . Further, the ratio of the pitch circle diameter PCDc of the balls 7c in the first row to the pitch circle diameter PCDd of the balls 7d in the second row is 98 to 70% from the viewpoint of achieving both a reduction in dynamic torque and securing a rolling fatigue life. Fit in range. If PCDc / PCDd exceeds 98%, the difference between PCDc and PCDd in both rows becomes too small, the height of the shoulders 13b and 15a cannot be secured, and the balls 7c are used when the ball bearing 1b is used. , 7d makes it easy for the rolling surfaces of the balls 7c, 7d to ride on the edges of the shoulders 13b, 15a. On the other hand, if it is less than 70%, the difference between PCDc and PCDd in both rows becomes excessive, and the thickness of the minimum thickness portion of the inner ring 6b and the height of the engaging portion 16a are ensured. Things get harder.

Further, the ratio D _S / D _L of the ball diameters of the first and second rows is 0.5 or more and less than 1 (0.5 ≦ D _S / D _L <1). When this ratio D _S / D _L is 1 or more, the effect of the present invention cannot be obtained. Conversely, if “D _S / D _L <1”, the effect of the present invention can be obtained to some extent. On the contrary, when the ratio D _{S /} D _L is less than 0.5, the ratio between the maximum inside diameter and minimum inside diameter of the inner peripheral surface of the outer ring 5b, and maximum outside of the outer peripheral surface of the inner ring 6b The ratio between the diameter and the minimum outer diameter is increased. As a result, for example, when the outer ring 5b and the inner ring 6b are manufactured by cold forging, the so-called two-piece cutting of the outer ring 5b and the inner ring 6b from a single material cannot be performed, and the manufacturing cost increases.
In any case, the thickness of each part of the outer ring 5b and the inner ring 6b (especially on the extension line of the load acting line and its vicinity) is sufficiently sufficient (for example, the radial direction of the outer ring and the inner ring at the deepest part of each track). The thickness of the ball is at least 25% of the diameter of the ball in rolling contact with the track.

  The tandem angular ball bearing of the present invention can be used not only for the rotation support portion of the pinion shaft of the differential gear as shown in FIG. 3, but also for various rotation support portions. For example, as shown in FIG. 2, when the tandem angular ball bearings 5b and 5b of the present invention are assembled in a front combination (DF combination) into the rotation support portions of the left and right output portions 17a and 17b of the differential gear, Balls 7c and 7c having a large ball diameter (and thus having a large rigidity and load capacity in each direction) are positioned on both ends of the rotation support portion. For this reason, even when a front combination which is disadvantageous in terms of securing moment rigidity is adopted, the moment rigidity of both the output portions 17a and 17b can be increased.

  Moreover, when implementing this invention, the dimension of each part etc. are suitably controlled in the range mentioned above in order to acquire the effect | action and effect mentioned above effectively. Further, even if the load applied to the tandem angular ball bearings is the ratio of the load of the balls in both rows, it is appropriately regulated from the viewpoint of effectively obtaining the above-mentioned effects.

1, 1a, 1b Ball bearing 2 Ball bearing 3 Pinion shaft 4 Pinion gear 5, 5a, 5b Outer ring 6, 6a, 6b Inner ring 7a, 7b, 7c, 7d Ball 8a, 8b Cage 9a, 9b, 9c, 9d Outer ring raceway 10a 10b, 10c, 10d Inner ring track 11a, 11b Pocket 12 Inner ring side assembly 13a, 13b Outer ring shoulder part 14a, 14b Counter bore part 15a, 15b Inner ring shoulder part 16a, 16b Engagement part 17a, 17b Output part

Claims (1)

An outer ring provided with two rows of outer ring raceways having different inner diameters on the inner peripheral surface, and an inner ring provided with two rows of inner ring raceways arranged on the outer peripheral surface concentrically with the outer ring and having different outer diameters on the outer peripheral surface And a plurality of balls provided between the outer ring raceways and the inner ring raceways so as to be able to roll in a state where a contact angle in the same direction is given between the rows, in each row. Tandem angular contact ball bearings with
Among the outer ring raceway and the inner ring raceway of the two rows, the ball diameter on the first row side having a small pitch circle diameter provided between the outer ring raceway and the inner ring raceway on the side having a smaller diameter, It is larger than the ball diameter on the second row side where the pitch circle diameter is large, which is provided between the outer ring raceway and the inner ring raceway on the larger side,
And by making the internal gap regarding the ball on the first row side smaller than the internal gap on the ball on the second row side,
In use, the load applied by the balls on the first row side is larger than the load applied by the balls on the second row side, and the balls on the first row side are loaded. A tandem angular contact ball bearing , wherein the load is 1.5 to 2.5 times the load applied by the ball on the second row side .

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