JP2014148993A - Double-row ball bearing and method of manufacturing the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To achieve a double-row ball bearing which can be constituted in a small size, has a high load capacity with respect to an axial load acting in both directions, and has excellent durability.SOLUTION: A difference between a pitch Pof both outer ring raceways 6e, 6f and a pitch Pof both inner ring raceways 7e, 7f is set to be 20 μm or less. A difference between inside diameters R, Rof groove bottoms of both the outer ring raceways 6e, 6f and a difference between outside diameters R, Rof groove bottoms of both the inner ring raceways 7e, 7f are each set to be 6 μm or less. Furthermore, a difference between radial gaps of ball bearings 11a, 11b in both rows constituted of both the outer ring raceways and both the inner ring raceways 6e, 6f, 7e, 7f and balls 4e, 4f is set to be 8 μm or less. By this constitution, an axial load applied on a double-row ball bearing 1b can be substantially-uniformly supported with the ball bearings 11a, 11b in both the rows.

Description

  The present invention relates to an improvement in a double-row ball bearing incorporated in a rotation support portion of various rotary machine devices such as an automobile transmission (including a transaxle) and a motor. Specifically, a double-row ball bearing having a large load capacity with respect to an axial load and having excellent durability and a manufacturing method thereof are realized.

  As a rolling bearing incorporated in a rotating support part of an automobile transmission or the like that generates a radial load and a thrust load in the power transmission part with the use of a gear such as a helical gear, it has a large load capacity and rotates. It is considered to use a double row ball bearing with a small resistance (dynamic torque). As a double row ball bearing that can be used in this case, for example, there is an angular type double row ball bearing as described in Patent Document 1 and Non-Patent Document 1. FIG. 7 shows the structure described in Patent Document 1 among them. The double-row ball bearing 1 includes one outer ring 2, a pair of inner rings 3a and 3b, a plurality of balls 4a and 4b, and a pair of cages 5a and 5b. Out of these, the outer ring 2 is provided with double-row outer ring raceways 6a and 6b on the inner peripheral surface, and is an integral type. The inner rings 3a and 3b are provided with inner ring raceways 7a and 7b on their outer peripheral surfaces, and are separate from each other. Further, a plurality of balls 4a and 4b are provided between the inner ring raceways 7a and 7b and the outer ring raceways 6a and 6b, respectively, so that they can roll freely. Further, both the cages 5a and 5b hold the balls 4a and 4b in both rows so as to roll freely. In this state, a preload is applied to each of the balls 4a and 4b as necessary, together with the contact angles α and β of the rear combination (DB) type. Such a double-row ball bearing 1 having a conventional structure can support a thrust load in both directions in addition to a radial load. Although not shown, the non-patent document 1 discloses a double-row ball in which the contact angle of the front combination (DF) type or the parallel (DT) type is given as a contact angle to be given to both rows of balls. The structure of the bearing is described.

In the first example of the conventional structure shown in FIG. 7, the balls 4a and 4b are provided in two rows, but the axial load is supported by the balls 4a and 4a (or one of the rows). 4b, 4b) only. Therefore, it is difficult to achieve both securing miniaturization and axial load capacity, a limited installation space, yet large (e.g., basic dynamic load rating C of more than 40% _or) under such conditions the axial load is applied If it is used, it may not always be possible to ensure sufficient durability.

  If a double row ball bearing provided with a parallel (DT) type contact angle as described in Non-Patent Document 1 is used, it is sufficient even when an axial load is supported by a relatively small double row ball bearing. It is easy to ensure the durability. However, in the case of the DT type structure, only an axial load in a predetermined direction can be supported (an axial load in the reverse direction cannot be supported). On the other hand, for example, in the case of a rotation support part of an automobile transmission, the acting direction of the axial load is reversed during acceleration and constant speed traveling and during deceleration (when the engine brake is activated). There are cases where it cannot be adopted.

  In view of such circumstances, it is conceivable to use a deep groove type double-row ball bearing that is neutral without an axial load and does not have a contact angle. Deep groove type double-row ball bearings support this axial load when an axial load is applied to each ball arranged in a double row with a contact angle in the direction corresponding to the direction of the action of this axial load. . In this case, if each ball arranged in a double row supports this axial load, the axial load capacity can be sufficiently secured and the durability of the double row ball bearing can be sufficiently secured.

  Examples of the deep groove type double row ball bearing that can be used for the rotation support portion as described above include those described in Patent Documents 2 to 4. 8-9 has shown the double row ball bearing 1a described in patent document 2 among these. This double row ball bearing 1a is formed on the inner peripheral surface of the outer ring 2a, each of which is a deep groove type outer ring raceway 6c, 6d, and formed on the outer peripheral surface of the inner ring 3c, each of which is a deep groove type inner ring raceway 7c, 7d, a plurality of balls 4c and 4d are arranged in both rows so as to be held by the cages 5c and 5d, respectively. In order to incorporate the balls 4c and 4d between the outer ring raceways 6c and 6d and the inner ring raceways 7c and 7d, the outer ring 2a and the inner ring 3c are eccentric as shown in FIG. In this state, a necessary number of balls 4c, 4c (4d, 4d) are assembled between the outer ring raceways 6c, 6d and the inner ring raceways 7c, 7d in a state of sticking to each other. 4c, 4c (4d, 4d) are equally distributed in the circumferential direction.

  In order to secure the durability of the double row ball bearing 1a by incorporating the deep groove type double row ball bearing 1a as described above into a rotation support portion to which an axial load is applied, the balls 4c, 4d of the both rows are arranged. Therefore, it is necessary to support the axial load almost uniformly. Only one row of balls 4c, 4c (4d, 4d) supports the axial load, or the proportion of balls 4c, 4c (4d, 4d) of either row supports this axial load is extremely high. When the number is increased, the rolling fatigue life of the balls 4c, 4c (4d, 4d) in one row is shortened, and the durability of the double row ball bearing 1a cannot be ensured.

  In Patent Document 3, a deep groove type double row ball bearing used as a turning wheel for rotating a wind turbine for wind power generation, the difference between the pitch of both outer ring raceways and the pitch of both inner ring raceways is set to 5 to 50 μm. For this reason, it is described that the outer ring raceway and the inner ring raceway are simultaneously finished. In a double row ball bearing, in order to support the axial load evenly with the balls in both rows, it is important to keep the difference between the pitch of both outer ring raceways and the pitch of both inner ring raceways small. It cannot be supported evenly. Therefore, only the technique described in Patent Document 3 cannot achieve both downsizing and ensuring of durability of the double-row ball bearing in an application as intended by the present invention.

  Further, in Patent Document 4, before assembling, the difference between the inner diameter of the outer ring raceway, the outer diameter of the inner ring raceway, the diameter of the ball with respect to each set reference dimension is obtained, and according to each difference, the outer ring, the inner ring, It describes a method for assembling a double row ball bearing in which the radial clearance of the double row ball bearing is set to a predetermined value by selectively combining the balls. However, the invention described in Patent Document 4 is intended only to set the radial clearance of the double row ball bearing to a predetermined value, and intended to support the axial load evenly by both rows of balls. is not. Therefore, even with the technique described in Patent Document 3 alone, it is impossible to achieve both downsizing of the double row ball bearing and ensuring of durability in an application intended by the present invention. Non-Patent Document 2 describes the relationship between the radial gap and the axial gap of the ball bearing.

JP 2011-252521 A JP 2009-41756 A International Publication No. 2009/147865 Pamphlet Japanese Patent Publication No. 63-649

“NSK Technical Report”, NSK Ltd., 2004, No. 728e E-3, p. 15, p. 143 “NSK Technical Report”, NSK Ltd., 2004, No. 728e E-3, p. 96

  In view of the circumstances as described above, the present invention was invented to realize a double row ball bearing that can be configured in a small size, has a large load capacity with respect to an axial load acting in both directions, and has excellent durability. .

A double-row ball bearing that is an object of the present invention includes an outer ring, an inner ring, and a plurality of balls.
Of these, the outer ring is provided with double-row outer ring raceways on the inner peripheral surface.
The inner ring is provided with double-row inner ring raceways on the outer peripheral surface.
Further, a plurality of balls are provided between the inner ring raceways and the outer ring raceways for each row so as to roll freely. The pitch circle diameter (PCD) of the balls in both rows is Are almost equal to each other. That is, the pitch circle diameters (PCD) of the balls in both rows are only slightly different based on the difference between the outer diameter of the inner ring raceways and the inner diameter of the outer ring raceways, and are basically the same. is there.

In particular, in the double row ball bearing of the present invention as set forth in claim 1, the outer ring raceways are each a deep groove type, and the shape and dimensions of the bus bar are substantially equal to each other (the bus bar shape is an arc shape). The difference in radius of curvature is 5 μm or less).
Each of the inner ring raceways is a deep groove type, and the shape and dimensions of the bus bar are substantially equal to each other (the bus bar shape is an arc shape and the difference in curvature radius is 5 μm or less).
Furthermore, each said ball | bowl has a substantially equal diameter mutually. That is, among these balls, the difference in diameter between the largest and smallest balls is within 4 μm (preferably within 2 μm).
The difference between the pitch of the outer ring raceways (interval in the axial direction between the groove bottoms) and the pitch of the inner ring raceways is suppressed to 20 μm or less, more preferably 10 μm or less.
Further, the difference between the inner diameters of the groove bottom portions of both outer ring raceways and the difference between the outer diameters of the groove bottom portions of both inner ring raceways are suppressed to 6 μm or less.
Furthermore, the difference between the radial clearances of the two rows of ball bearings each constituted by the outer ring and inner ring raceways and the balls is set to 8 μm or less, more preferably 6 μm or less.

The main point of the double row ball bearing of the present invention configured as described above is to reduce the difference between the total axial gaps described below, so that the axial loads (shared axial loads) supported by both rows of ball bearings are shared. the difference is sufficiently small (e.g., the suppressed to less than several% of the basic dynamic load rating C _or double row ball bearings) in point.
Here, regarding the double-row ball bearing of the present invention, the total axial gap difference is the difference (Δ1) between the pitch of the outer ring raceways and the pitch of the inner ring raceways, and the axial gap of the ball bearings of both rows. It is an amount represented by the sum (Δ1 + Δ2) of the difference (Δ2) between them. In addition, the axial clearance for each of the ball bearings in both rows is the radial clearance, the radius of curvature of the bus bar shape of the outer ring raceway, the radius of curvature of the bus race shape of the inner ring raceway, This is a calculated value obtained by the formula described in Non-Patent Document 2 based on the diameter.
Therefore, in the double row ball bearing of the present invention, the difference between the pitch of the outer ring raceways and the pitch of the inner ring raceways, the difference between the inner diameters of the groove bottom portions of the outer ring raceways, and the groove bottom portions of the inner ring raceways. The difference between the outer diameters of the ball bearings and the difference between the radial gaps of the ball bearings of the two rows are within the ranges as described above, thereby reducing the total axial gap difference, thereby reducing the balls of the rows of the two rows. The shared axial load of the bearing is made almost equal.

When implementing the double row ball bearing of the present invention as described above, the outer ring, the inner ring, and the balls are all made of a hard iron-based alloy. As this iron-based alloy, high carbon chromium bearing steel (SUJ) specified in JIS G 4805 is the most common, but stainless steel (SUS), carbon can be used as long as the required hardness can be secured. Tool steel (SK), high speed tool steel (SKH), etc. can also be used. However, ceramics have a much larger elastic coefficient than these iron-based alloys, and thus the above values are not applied as they are, and are not suitable as materials for carrying out the present invention.
For example, when the outer ring and the inner ring and the balls are made of an iron-based alloy such as SUJ2, the diameter of the balls is set to 5 to 20 mm as in the invention described in claim 2, for example. The pitch between both outer ring raceways and both inner ring raceways is set to 35 to 130 mm.

Moreover, when implementing the double row ball bearing of this invention, Preferably, each ball which comprises the ball bearing of the said both row | line | column is each made from a synthetic resin, and an annular shape like the invention described in Claim 3. The rim portion is rotatably held by a pair of crown-shaped cages provided with pockets on one side surface in the axial direction. And the rim | limb part of these both crown type holder | retainers is arrange | positioned on the opposite side of an axial direction on both sides of the said each ball | bowl.
Alternatively, as in the invention described in claim 4, a pair of axial end openings of the bearing inner space in which the balls are installed are closed between the inner peripheral surface of the outer ring and the outer peripheral surface of the inner ring. Provide a seal ring.
When the invention of such a double row ball bearing described in claim 4 is carried out, preferably, the flow direction of the lubricating oil and the acting direction of the axial load are regulated as in the invention described in claim 5. That is, when the double row ball bearing is used, the acting direction of the axial load applied between the outer ring and the inner ring is a predetermined direction in the majority of the usage time, and the rotation support part is a part of the two seal rings. One seal ring whose peripheral edge tends to be lifted from the mating surface based on the axial load acting in the predetermined direction is incorporated in a state where it is installed on the downstream side with respect to the flow direction of the lubricating oil flowing through the rotation support portion.

Further, the invention of the method for manufacturing a double row ball bearing according to claim 6 includes an outer ring having a double row outer ring raceway on the inner peripheral surface, an inner ring having a double row inner ring raceway on the outer peripheral surface, Between the outer ring raceway and the inner ring raceways, a plurality of balls are provided so as to be able to roll in each row, and the outer ring raceways are deep groove type, and the shape and dimensions of the bus bar are substantially the same. Equally, the inner ring raceways are each a deep groove type, the shape and dimensions of the bus bar are substantially equal to each other, and the balls have a diameter substantially equal to each other.
The invention of the manufacturing method of such a double row ball bearing of the present invention is such that both the outer ring raceways are simultaneously polished by, for example, a polishing apparatus equipped with a rotary dresser, and the difference between the inner diameters of the groove bottom portions of these outer ring raceways is determined. The inner ring raceways are simultaneously polished to 6 μm or less by, for example, a polishing device equipped with a rotary dresser, and the difference between the outer diameters of the groove bottoms of these inner ring raceways is suppressed to 6 μm or less. And the difference between the pitches of the two inner ring raceways are suppressed to 20 μm or less (preferably 10 μm or less).
Then, one outer ring and one inner ring selected from each of the plurality of outer rings and inner rings are combined. Next, the combination direction of the outer ring and the inner ring is regulated so that the radial gap difference between the two rows is reduced according to the inner diameter of the groove bottom of the outer ring raceway and the outer diameter of the groove bottom part of the inner raceway. In addition, the outer ring and the inner ring are combined in a state in which the balls having substantially the same diameter between these tracks (for example, the difference in diameter between each track is suppressed to 2 to 4 μm or less) are arranged in a freely rolling manner. The difference between the radial clearances of the ball bearings in both rows each constituted by the outer ring and inner ring raceways and the balls is set to 8 μm or less (preferably 6 μm or less).

According to the present invention configured as described above, it is possible to realize a double row ball bearing that can be configured in a small size, has a large load capacity with respect to an axial load acting in both directions, and has excellent durability.
That is, in the case of the double row ball bearing of the present invention, the difference between the pitch of the pair of outer ring raceways and the pitch of the pair of inner ring raceways, the difference between the inner diameters of the groove bottoms of these outer ring raceways, The difference between the outer diameters of the groove bottoms of the inner ring raceway and the radial clearance of the ball bearings in both rows are small values, but there are manufacturing problems such as a high defect rate. The value is kept at a level that does not invite. Then, by appropriately regulating the combination state of the outer ring, the inner ring, and each ball, the total axial gap difference of the double row ball bearing {= the difference between the pitch of the outer ring raceway and the pitch of the inner ring raceway (Δ 1) and the sum (Δ1 + Δ2)} of the difference (Δ2) between the axial gaps of the ball bearings in both rows are kept small. As a result, when an axial load is applied between the outer ring and the inner ring, the ball bearings in both rows support the axial load almost evenly. In other words, the difference between the share axial load between the ball bearings of both rows can be sufficiently reduced (for example be suppressed to less than several% of the basic dynamic load rating C _or of the double row ball bearing). For this reason, it is possible to suppress an increase in the surface pressure of the rolling contact portion related to the ball bearings in both rows, and to increase the load capacity related to the axial load without using a large-diameter outer ring, inner ring, or each ball. it can. For this reason, durability can be secured while suppressing an increase in size of the double row ball bearing.

  Further, as in the third aspect of the invention, it is advantageous in terms of realizing stable high-speed rotation if each ball constituting the both-row ball bearings is held by a synthetic resin crown-shaped cage. . First, the use of a synthetic resin-made crown-type cage that is lighter than metal is advantageous in terms of suppressing the inertial mass of the cage and realizing high-speed rotation. Also, by using a cage made of synthetic resin that is superior in self-lubrication compared to metal, the time until serious damage such as seizure occurs even when operating under dilute or low lubrication. This is advantageous from the standpoint that the length can be increased. In addition, the rim | limb part of a pair of crown type holder | retainer holding each said ball | bowl is arrange | positioned on the opposite side of an axial direction on both sides of these each ball | bowl. The reason for this is to improve the workability of assembling the double crown cage to the double row ball bearing of the present invention. Further, by not arranging the rim between the balls in both rows, it is possible to prevent the pitch between the outer ring raceways and the inner ring raceways from becoming excessive. For this reason, the axial rigidity of both side portions of both outer ring raceways and the both side portions can be made equal to each other with limited axial dimensions, and the axial loads in both directions are evenly supported by the ball bearings in both rows. It becomes advantageous from a surface.

Alternatively, as in the invention described in claim 4, if both axial end openings of the bearing internal space are closed by a pair of seal rings, the gear of the gear mixed in the lubricating oil circulating in the casing constituting the transmission is mixed. It is possible to suppress the deterioration of the durability of the double row ball bearing by suppressing the entry of foreign matter such as wear powder into the bearing internal space.
In this case, for example, as in the invention described in claim 5, if one seal ring whose peripheral edge tends to lift from the mating surface based on the axial load is installed on the downstream side in the flow direction of the lubricating oil, Regardless of the deterioration of the sealing performance associated with lifting, foreign matter intrusion into the internal space can be suppressed.
Furthermore, according to the manufacturing method of the double row ball bearing of the present invention described in claim 6, the processing accuracy of the outer ring, the inner ring, and each ball, each of which is a constituent member of the double row ball bearing, is manufactured. Even if the cost is not so high as to increase the cost, a double row ball bearing can be obtained that exhibits the above-described excellent actions and effects. In other words, the manufacturing cost of the double row ball bearing of the present invention can be suppressed.

The fragmentary sectional view of the double row ball bearing which shows one example of embodiment of this invention. The flowchart which shows the assembly procedure of the structure of this invention. The diagram which shows the relationship between the total axial gap difference and the difference of the axial load which both rows support. The figure similar to FIG. 1 for demonstrating the effect of this invention. The bar graph which shows the result of the experiment conducted in order to confirm the effect of this invention. Similarly, a graph showing the relationship between the total axial gap difference and the service life. Sectional drawing which shows the 1st example of a conventional structure. The partial cut perspective view which shows the 2nd example. Similarly sectional drawing. Sectional drawing which shows the state which incorporates a ball between an outer ring track and an inner ring track.

An example of an embodiment of the present invention will be described with reference to FIGS. As shown in FIG. 1, the double row ball bearing 1b of this example includes an outer ring 2b, an inner ring 3d, a plurality of balls 4e and 4f, a pair of cages 5e and 5f, and a pair of seal rings 8a. , 8b.
Out of these, the outer ring 2b is made of high-carbon chromium bearing steel class 2 (SUJ2), and double row outer ring raceways 6e and 6f are provided on the inner peripheral surface. These outer ring raceways 6e and 6f are each a deep groove type, and the shape and dimensions of the bus bar are substantially equal to each other. That is, these outer ring raceways 6e, 6f, the generatrix shape in arc shape, the both outer ring raceways 6e, among 6f each other, the difference in the radius of curvature R ₆ of the circular arc is 5μm or less. It should be noted that the radius of curvature R ₆ of the generatrix shape of these outer ring raceways 6e, 6f is slightly larger than ½ of the diameter D ₄ of each of the balls 4e, 4f {eg, R ₆ = (0.515-0 .55) D ₄ }. In addition, the inner diameters of the three outer ring shoulder portions 9a, 9b, 9c existing at the positions sandwiching the outer ring raceways 6e, 6f from both sides in the width direction on the inner peripheral surface of the outer ring 2b are the same, and the shaft The directional dimensions are almost the same as each other. Accordingly, the axial rigidity of each of the outer ring shoulder portions 9a, 9b, 9c is the same. Furthermore, the outer ring raceways 6e, the groove bottom of the inner diameter _R 6e of _6f, the difference between _{R 6f,} the 6 [mu] m or less _{(| R 6e -R 6f | ≦} 6μm) is suppressed.

The inner ring 3d is made of SUJ2 and has double-row inner ring raceways 7e and 7f on the outer peripheral surface. These inner ring raceways 7e and 7f are each a deep groove type, and the shape and dimensions of the bus bar are substantially equal to each other. In other words, these two inner ring raceways 7e, 7f, the shape of the generating line is a circular arc shape, these two inner ring raceways 7e, among 7f each other, the difference in the curvature radius R ₇ of the circular arc is 5μm or less. Incidentally, these two inner ring raceways 7e, 7f, the respective balls 4e, 4f slightly larger than half the diameter _{D 4} of, further, the outer ring raceways 6e, is less than the radius of curvature _{R 6} relates 6f {e.g. , R ₇ = (0.505-0.54) D ₄ }. In addition, the outer diameters of the three inner ring shoulder portions 10a, 10b, and 10c existing at positions sandwiching the inner ring raceways 7e and 7f from both sides in the width direction on the outer peripheral surface of the inner ring 3d are the same as each other. The directional dimensions are almost the same as each other. Therefore, the axial rigidity of each of the inner ring shoulder portions 10a, 10b, and 10c is the same. Furthermore, the two inner ring raceway 7e, 7f groove bottom portion of the outer diameter _R 7e _of, even concerning the difference between _{R 7f,} to 6 [mu] m or less _{(| R 7e -R 7f | ≦} 6μm) is suppressed.

Further, the outer ring raceways 6e, the pitch _{P 6} of 6f, the two inner ring raceway 7e, the pitch _{P 7} of 7f are both of the respective balls 4e, 4f diameter _{D 4} of 1.1 to 1.5 Double {P ₆ , P ₇ = (1.1 to 1.5) D ₄ }. The difference between the pitches P ₆ and P ₇ is suppressed to 20 μm or less (| P ₆ −P ₇ | ≦ 20 μm), more preferably 10 μm or less (| P ₆ −P ₇ | ≦ 10 μm).

Further, the respective balls 4e, 4f has the same diameter D ₄ each other, the outer ring raceways 6e, 6f and the two inner ring raceway 7e, between 7f, by several same each other between the two rows to each other, Each of the cages 5e and 5f is provided so as to be freely rollable. The difference between the radial gaps of the ball bearings 11a and 11b in both rows is 8 μm or less, more preferably 6 μm or less. In addition, the value of the radial clearance of the ball bearings 11a and 11b in both rows is regulated from the viewpoint of securing required performance such as axial rigidity.

  Further, both the cages 5e and 5f are crown-shaped cages formed by injection molding a synthetic resin having self-lubricating properties such as polyamide resin, and each of the cages 5e and 5f includes a plurality of cages on one side surface in the axial direction of the annular rim portion 12. The elastic claw portions 13 are provided, and the portions surrounded on three sides by the axially one side surface of the rim portion 12 and the elastic claw portions 13 adjacent to each other in the circumferential direction are defined as pockets 14. Balls 4e and 4f are held so as to roll freely. In this state, the rim parts 12 and 12 of both the cages 5e and 5f are arranged on the opposite side in the axial direction with the balls 4e and 4f interposed therebetween. With this configuration, each of the shoulder portions 9a, 9b, 4b, 4f, which form the ball bearings 11a, 11b in both rows, is appropriately brought close to each other without excessively increasing the width of the outer ring 2b and the inner ring 3d. The axial dimensions of 9b, 9c, 10a, 10b, and 10c are sufficiently secured, and the axial dimensions of the shoulders 9a, 9b, 9c, 10a, 10b, and 10c are made substantially uniform.

  Further, the seal rings 8a and 8b are formed in an annular shape as a whole and are formed by reinforcing an elastic material made of elastomer such as rubber with a cored bar. Then, the outer peripheral edge portions are engaged with the engaging grooves 15a and 15b formed on the inner peripheral surfaces of the both ends of the outer ring 2b, and the seal lips 16a and 16b provided on the inner peripheral edge portions are connected to the inner ring 3d. The stepped surfaces 17a and 17b formed on the outer peripheral surfaces of both ends are brought into sliding contact with the entire periphery. With this configuration, both axial ends of the bearing internal space 18 in which the balls 4e and 4f are installed are closed between the inner peripheral surface of the outer ring 2b and the outer peripheral surface of the inner ring 3d.

As described above, the double row ball bearing 1b of this example, in which the dimensional difference between the respective parts is suppressed to a predetermined value or less, selects and combines the outer ring 2b, the inner ring 3d, and the balls 4e, 4f, each of which is formed in large numbers. Consists of things. As described above, the difference between the inner diameters R _6e and R _6f of the groove bottoms of the double row outer ring raceways 6e and 6f on the inner circumferential surface of the outer ring 2b, and the double row on the outer circumferential surface of the inner ring 3d. inner raceway 7e, 7f groove bottom portion of the outer diameter _{R 7e,} the difference between _{R 7f} are all kept suppressed to 6μm or less. Similarly, as described above, the difference between the pitch P ₆ between the outer ring raceways 6e and 6f and the pitch P ₇ between the inner ring raceways 7e and 7f is also 20 μm or less (preferably 10 μm or less). Keep it down. For this purpose, both the outer ring raceways 6e and 6f are simultaneously polished by a polishing apparatus equipped with a rotary dresser, and the inner ring raceways 7e and 7f are similarly polished. For each of the balls 4e and 4f, the diameters of the balls 4e and 4f are measured in advance with a gauge, and the diameter difference between them is 2 to 4 μm or less. Prepare for all columns.

If the outer ring 2b, the inner ring 3d, and the balls 4e and 4f, each satisfying the above conditions, are prepared, as shown in FIG. 2, first, in step 1 (S1), a large number of outer rings 2b are produced. Any one outer ring 2b (or inner ring 3d) is selected from (or inner ring 3d).
Next, in step 2 (S2), the inner diameters R _6e and R _6f of the outer ring raceways 6e and 6f of the selected outer ring 2b (or inner ring 3d) (or the outer diameters R _7e and R _{7f of the} inner ring raceways 7e and _7f ) are measured. To do.
Next, in Step 3 (S3), any one of the inner rings 3d (or outer rings 2b) is selected from among a large number of inner rings 3d (or outer rings 2b).
Next, in step 4 (S4), the outer diameters R _7e and R _7f of the inner ring raceways 7e and 7f of the selected inner ring 3d (or outer ring 2b) (or the inner diameters R _6e and R _{6f of the} outer ring raceways 6e and _6f ) are measured. To do.
Next, in step 5 (S5), the combination direction of the outer ring 2b and the inner ring 3d that can reduce the difference in radial clearance between the ball bearings 11a and 11b in both rows is determined.
Next, in Step 6, the radials of the ball bearings 11a and 11b of the double-row ball bearing 1b after completion based on the measured values in Steps 2 and 4, the predetermined ball diameter range, and the assembly direction determined in Step 5 are obtained. A clearance is obtained, and it is determined whether or not the radial clearance between the ball bearings 11a and 11b in both rows falls within a predetermined value range. As a result of this determination, when it is determined that the radial clearances of the ball bearings 11a, 11b in both rows cannot be accommodated within a predetermined value range, the process returns to the step 3, and another inner ring 3d (or outer ring 2b) is returned. In step 4 (S4), the outer diameters R _7e and R _7f of the inner ring raceways 7e and 7f of the newly selected inner ring 3d (or outer ring 2b) (or the inner diameters R _6e and R _{6f of the} outer ring raceways 6e and 6f) are selected. ) And the combination direction of the outer ring 2b and the inner ring 3d is determined in step 5, and then the process returns to step 6 again.
If it is determined in step 6 that the radial clearances of the ball bearings 11a and 11b in both rows are within a predetermined range, the combination direction determined in step 5 in the last step 7 (S7). Accordingly, the inner ring 3d is disposed on the inner diameter side of the outer ring 2b, and as shown in FIG. 10 described above, between the outer ring raceways 6e, 6f and the inner ring raceways 7e, 7f, Incorporate balls 4e, 4f.
By the above operation, the double row ball bearing 1b in which the dimensions of the respective parts are regulated as described above can be obtained. Note that the two inner ring raceway 7e, the pitch P ₇ of 7f, the outer ring raceways 6e, the pitch P ₆ of 6f is precisely regulated by a polishing apparatus having a rotary dresser, respectively. For this reason, the pitch difference is always within the range of 20 μm for any combination of the outer ring 2b and the inner ring 3d. Therefore, it is not necessary to set a determination step for this pitch difference.

  When the double-row ball bearing 1b having the above-described configuration assembled as described above is incorporated in, for example, a rotation support portion of a power transmission shaft of an automobile transmission, the outer ring 2b is provided in, for example, a casing. The inner ring 3d is fitted and fixed to the power transmission shaft. When the automobile transmission is operated in this state, a radial load and an axial load are applied to the double row ball bearing 1b by a gear reaction force generated at the meshing portion of the helical gears constituting the power transmission device. Here, the total axial gap difference of the double row ball bearing 1b {= the difference (Δ1) between the pitch of the outer ring raceways 6e, 6f and the pitch of the inner ring raceways 7e, 7e, and the ball bearings of the double row When the sum (Δ1 + Δ2)} of the difference (Δ2) between the axial gaps 11a and 11b is zero, there is a difference between the shared axial loads of the ball bearings 11a and 11b in both rows. It never happens. On the other hand, when the difference between the total axial gaps is not zero, a difference occurs between the shared axial loads of the ball bearings 11a and 11b in both rows. And the difference of this shared axial load becomes large, so that the said total axial gap difference becomes large. FIG. 3 shows this relationship.

In this regard, in the case of the double row ball bearing 1b of this example, the total axial gap difference is suppressed to be small by suppressing the dimensional difference between the respective parts slightly as described above. ball bearings 11a, can sufficiently reduce the difference in shared axial load between the 11b (e.g. suppressed to less than several% of the basic dynamic load rating C _or of the double row ball bearing). For this reason, the surface pressure rise of the rolling contact part regarding these ball bearings 11a and 11b for both rows can be suppressed.

Further, in the case of the double row ball bearing 1b of this example, with respect to the radial load, the axial load does not cause a difference in the ratio of the support of the ball bearings 11a and 11b in both rows. Therefore, if the load ratio of the axial load is kept small, the load ratio of the radial load can be kept low enough to cause no problem.
As a result, the load relating to the axial load can be achieved without using the outer ring 2b, the inner ring 3d, and further each ball 4e, 4f that constitute the double row ball bearing 1b. The capacity can be increased. For this reason, the durability of the double row ball bearing 1b can be ensured while suppressing an increase in size.

  When an axial load F acts between the outer ring 2b and the inner ring 3d in the direction indicated by the white arrow in FIG. 4, the outer ring 2b and the inner ring 3d are axially moved based on elastic deformation of each part. Relative displacement. With this relative displacement, the surface pressure of the sliding contact portion between the seal lip 16a of the right seal ring 8a and the step surface 17a of the inner ring 3d increases in FIGS. The surface pressure of the sliding contact portion between the seal lip 16b of the seal ring 8b and the stepped surface 17b of the inner ring 3d is reduced or lost. Therefore, in such a state, in order to suppress entry of foreign matter into the bearing internal space 18, the relationship between the direction of action of the axial load F and the direction of flow of the lubricating oil flowing through the casing may be regulated. preferable.

  For example, when the double-row ball bearing 1b is incorporated in the rotation support portion of the power transmission shaft of an automobile transmission, the double-row ball bearing 1b is shown in FIG. When the axial load F is applied in the direction indicated by the thick bold arrow, the flow direction of the lubricating oil is changed from the right to the left in FIG. As a result, foreign matter such as gear wear powder from the gap between the seal lip 16b and the stepped surface 17b, which tends to float on the basis of the axial load acting over the majority of the usage time, is generated in the bearing inner space 18. It can be suppressed to enter.

An experiment conducted for confirming the effect of the present invention will be described. In the experiment, five types of samples were prepared: two types of samples belonging to the technical scope of the present invention (Examples 1 and 2), and three types of samples outside the technical scope of the present invention (Comparative Examples 1 to 3). In each case, an experiment for durability was conducted. The experimental conditions are as follows.
"Common conditions for each sample"
Outer diameter of outer ring 2b: 55mm
Inner ring 3d inner diameter: 30 mm
Diameter D ₄ of each ball 4e, 4f: 7.144mm
Radius of curvature R ₆ , R ₇ : 3.715 mm (same for each other)
"Different conditions for each sample"
In the case of Example 1 Difference between pitches P ₆ and P _{7 of the} outer ring and inner ring raceways: 10 μm
Difference between inner diameters of both outer ring raceways 6e and 6f: 0 μm
Difference between outer diameters of both inner ring raceways 7e and 7f: 0 μm
Difference between radial clearances of both ball bearings 11a and 11b: 0 μm
In the case of Example 2 Difference between pitches P ₆ and P _{7 of the} outer ring and inner ring raceways: 20 μm
Difference between inner diameters of both outer ring raceways 6e and 6f: 6 μm
Difference between outer diameters of both inner ring raceways 7e and 7f: 6 μm
Difference between radial clearances of both ball bearings 11a, 11b: 8 μm
In the case of Comparative Example 1, the difference between the outer ring and inner ring pitches P ₆ and P ₇ : 30 μm
Difference between inner diameters of both outer ring raceways 6e and 6f: 6 μm
Difference between outer diameters of both inner ring raceways 7e and 7f: 6 μm
Difference between radial clearances of both ball bearings 11a, 11b: 8 μm
In the case of Comparative Example 2, the difference between the outer ring and inner ring pitches P ₆ and P ₇ : 20 μm
Difference between inner diameters of both outer ring raceways 6e and 6f: 8 μm
Difference between outer diameters of both inner ring raceways 7e and 7f: 8 μm
Difference between radial clearances of both ball bearings 11a, 11b: 8 μm
In the case of Comparative Example 3 Difference between pitches P ₆ and P _{7 of the} outer ring and inner ring raceways: 20 μm
Difference between inner diameters of both outer ring raceways 6e and 6f: 6 μm
Difference between outer diameters of both inner ring raceways 7e and 7f: 6 μm
Difference between radial clearances of both ball bearings 11a and 11b: 10 μm

5 types of double row ball bearings each having the above specifications are operated under the same conditions {inner ring rotation, rotational speed = 5000 min ⁻¹ , radial load = 0, axial load = 6000 N (= 0.35 C _or )}. The rolling fatigue life of each was determined. The result is shown in FIG. In FIG. 5, the rolling fatigue life of Example 1 is set to 1, and the rolling fatigue life of other samples is shown as a ratio with respect to Example 1. As can be seen from FIG. 5 in which such an experimental result is described, the difference between the pitches P ₆ and P ₇ of the outer ring and inner ring raceways, the difference between the inner diameters of both outer ring raceways 6e and 6f, the inner ring raceways 7e and 7f If any one of the difference between the outer diameters and the difference between the radial gaps of the ball bearings 11a and 11b is larger than the value defined by the present invention, the durability of the double row ball bearing is greatly impaired. In other words, according to the present invention in which each of the above differences is regulated to a predetermined value or less, the durability of the double row ball bearing can be sufficiently secured.

  Furthermore, FIG. 6 shows the relationship between the values obtained by converting the dimensional differences of the samples into total axial gap differences and the rolling fatigue life of the samples. As apparent from a comparison between Example 2 and Comparative Example 3 in which only the difference in gap is different, the rolling fatigue life can be greatly improved by changing the difference in radial gap from 10 μm to 8 μm.

  The present invention is not limited to the automobile transmission and motor rotation support portions of the present invention, and can be applied to rotation support portions of various rotary machine devices in which an axial load is applied in addition to a radial load during operation.

1, 1a, 1b Double row ball bearing 2, 2a, 2b Outer ring 3a, 3b, 3c, 3d Inner ring 4a, 4b, 4c, 4d, 4e, 4f Ball 5a, 5b, 5c, 5d, 5e, 5f Cage 6a, 6b, 6c, 6d, 6e, 6f Outer ring raceway 7a, 7b, 7c, 7d, 7e, 7f Inner ring raceway 8a, 8b Seal ring 9a, 9b, 9c Outer ring shoulder 10a, 10b, 10c Inner ring shoulder 11a, 11b Ball bearing 12 Rim part 13 Elastic claw part 14 Pocket 15a, 15b Locking groove 16a, 16b Seal lip 17a, 17b Step surface 18 Bearing internal space

Claims (6)

  An outer ring having a double-row outer ring raceway on the inner peripheral surface, an inner ring having a double-row inner ring raceway on the outer peripheral surface, and a plurality of each for both rows between these outer ring raceways and these inner ring raceways. In a double-row ball bearing provided with balls that are freely rollable, the outer ring raceways are deep groove types, and the shape and dimensions of the busbars are substantially equal to each other. Each is a deep groove type, and the shape and dimensions of the bus bar are substantially equal to each other, the balls have diameters substantially equal to each other, and the difference between the pitch of the outer ring raceways and the pitch of the inner ring raceways is The difference between the inner diameters of the groove bottom portions of these outer ring raceways and the difference between the outer diameters of the groove bottom portions of both inner ring raceways are both 6 μm or less. And the ball bearings in both rows composed of the balls. A double row ball bearing characterized in that a difference between radial gaps is 8 μm or less.   The outer ring, the inner ring, and the balls are made of an iron-based alloy, the diameter of the balls is 5 to 20 mm, and the pitch of the outer ring raceways and the inner ring raceways is 35 to 130 mm. The double row ball bearing described in 1.   Each ball constituting the ball bearings in both rows is made of a synthetic resin, and is held by a pair of crown type cages having a pocket on one side surface in the axial direction of an annular rim portion. The double-row ball bearing according to any one of claims 1 and 2, wherein the rim portions of both the crown type cages are arranged on the opposite side in the axial direction with the balls interposed therebetween.   A pair of seal rings are provided between the inner peripheral surface of the outer ring and the outer peripheral surface of the inner ring to close both axial openings of the bearing internal space where the balls are installed. The double row ball bearing described in any one of the above.   Axial load acting between the outer ring and the inner ring in the use state, in which the acting direction of the axial load applied to the predetermined direction in the majority of the use time acts in the predetermined direction of the seal rings. 5. One seal ring, whose peripheral edge tends to be lifted from the mating surface based on the load, is incorporated in a state of being installed downstream with respect to the flow direction of the lubricating oil flowing through the rotation support portion. The double row ball bearing described.   An outer ring having a double-row outer ring raceway on the inner peripheral surface, an inner ring having a double-row inner ring raceway on the outer peripheral surface, and a plurality of each for both rows between these outer ring raceways and these inner ring raceways. The outer ring raceways are each a deep groove type and the shape and dimensions of the busbars are substantially equal to each other, and the inner ring raceways are each a deep groove type and the shape and dimensions of the busbars are both A method of manufacturing a double row ball bearing, wherein the balls are substantially equal to each other, and the balls have substantially the same diameter, and the difference between the inner diameters of the groove bottoms of the outer ring raceways is 6 μm or less, And the difference between the outer diameters of the inner ring raceway is suppressed to 6 μm or less, and the difference between the pitch of the outer ring raceways and the pitch of the inner ring raceways is suppressed to 20 μm or less. One outer ring selected from A combination of the outer ring and the inner ring so that the radial gap difference between both rows is reduced according to the inner diameter of the groove bottom of the outer ring raceway and the outer diameter of the groove bottom of the inner ring raceway. The outer ring and the inner ring are combined in a state in which the balls having substantially the same diameter between the respective raceways are arranged so as to be capable of rolling, and the outer ring and the inner ring both raceways and the respective raceways are combined. A method for producing a double row ball bearing, wherein the difference between the radial gaps of the ball bearings in both rows composed of balls is 8 μm or less.

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