JP6500356B2 - Bearing mechanism - Google Patents

Description

  The present invention relates to a bearing mechanism for supporting a gear shaft having a gear at one end, such as a pinion shaft of a differential gear of a motor vehicle, for example.

  One of the means for improving the fuel consumption of automobiles is to reduce the rotational resistance of the drive system. Since a hypoid gear (curved bevel gear) is used for the pinion of the differential gear (final reduction gear) in the drive system, a large radial load and axial load are applied to the pinion shaft. Therefore, conical roller bearings with high rigidity are often used for the pinion shaft.

  However, although the tapered roller bearing has an advantage of high rigidity, it has problems of high rotational resistance and large loss torque. Therefore, Patent Document 1 proposes a configuration in which, of the two bearings attached to the pinion shaft, the pinion side is a double-row tandem angular ball bearing and the connection flange side is a single-row angular ball bearing. . According to Patent Document 1, it is described that the rotation torque can be reduced by the above configuration, and the fuel efficiency of the vehicle is improved.

Patent No. 4250952

  Angular contact ball bearings are known to be capable of receiving both radial and axial loads. However, even if double-row angular ball bearings are used, the reduction in rigidity is inevitable compared to tapered roller bearings. And since a very large load is applied to the pinion shaft, if it is attempted to have the same rigidity as the tapered roller bearing, the size of the bearing must be increased. When the size of the bearing is increased, there is a problem that the fuel consumption is reduced due to the increase in weight of the vehicle, and the degree of freedom in the layout of components is reduced in designing the vehicle.

  Then, an object of this invention is to propose the bearing mechanism which can suppress a rigid fall, reducing loss torque rather than the conventional conical roller bearing.

  In order to solve the above-mentioned subject, a typical composition of the present invention is a bearing mechanism which supports a gear shaft provided with a gear at one end, and the first bearing disposed in the vicinity of the gear and the gear shaft And a second bearing disposed on the opposite side of the gear, wherein the first bearing is a double-row angular contact ball bearing, wherein one row of the double-row is multipoint contact.

  According to the above configuration, the reduction in rigidity can be suppressed while the ball bearing is used, and therefore the loss torque can be reduced as compared with the tapered roller bearing without increasing in size from the tapered roller bearing. In addition, since the rigidity is high as compared with the case of using only a single-row or double-row angular ball bearing, it is possible to suppress the meshing point displacement amount of the gear to a small amount. The multipoint contact means that the ball and the raceway surface are in contact at three or more points.

  The second bearing is preferably a multi-point contact single-row angular contact ball bearing. The multi-point contact also in the second bearing makes it possible to reduce the amount of displacement of the connection on the opposite side of the gear (for example, the companion flange).

  The row of multi-point contact of the first bearing is four-point contact, and the contact angles θ1 and θ2 between the raceway surface and the ball are both on one side of a line perpendicular to the gear axis, and the contact angle θ1 is 5 ° The contact angle θ2 is preferably in the range of 40 ° to 50 °.

  According to the above configuration, since the radial load and the axial load can be received in a well-balanced manner, high rigidity can be exhibited with respect to the load in either direction.

  According to the bearing mechanism of the present invention, it is possible to suppress the reduction in rigidity while reducing the loss torque as compared with the conventional tapered roller bearing.

It is a fragmentary sectional view explaining a bearing mechanism concerning an embodiment. It is a figure explaining the detailed composition of a bearing. It is a figure explaining the composition of an example. It is a figure explaining the composition of a comparative example. It is a figure which contrasts an example and a comparative example.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the accompanying drawings. The dimensions, materials, and other specific numerical values and the like shown in the embodiments are merely examples for facilitating the understanding of the invention, and the invention is not limited except as otherwise described. In the specification and the drawings, elements having substantially the same functions and configurations will be denoted by the same reference numerals to omit repeated descriptions, and elements not directly related to the present invention may be illustrated or described. Omit.

  FIG. 1 is a partial cross-sectional view for explaining a bearing mechanism according to the embodiment. In FIG. 1, the ring gear 102 of the differential gear and the pinion gear 104 mesh with each other. The pinion shaft 106 (gear shaft) includes a pinion gear 104 at one end, and a flange 108 (also referred to as a companion flange or connection flange) at the other end. The flange 108 is a component for connecting to a propeller shaft (not shown).

  The pinion shaft 106 is rotatably supported by the first bearing 200 and the second bearing 300. The first bearing 200 is disposed in the vicinity of the pinion gear 104 with its front facing the gear side. The second bearing 300 is disposed in the vicinity of the flange 108 (on the side of the pinion shaft 106 opposite to the pinion gear 104) with the front facing the flange 108 side. The bearing mechanism according to the present invention is constituted by the first bearing 200 and the second bearing 300.

  FIG. 2 is a view for explaining the detailed configuration of the bearing. As a feature of the present invention, the first bearing 200 is a double row angular contact ball bearing, and one row of the double rows is a multipoint contact. In the example of FIG. 2, among the double rows of the first bearing 200, the row on the gear side is in four-point contact.

  In the first bearing 200, balls 206 a and 206 b are disposed in a double row of raceways formed between the outer ring 202 and the inner ring 204. The contact angles .theta.1 and .theta.2 between the balls 206a of the gear side row and the raceway surface are both on one side of a line (shown by a broken line in FIG. 2) perpendicular to the pinion shaft 106 (gear shaft). The contact angle θ1 of the ball 206a is preferably in the range of 5 ° to 15 °, and the contact angle θ2 is in the range of 40 ° to 50 °. As an example, the contact angle θ1 can be 5 °, and the contact angle θ2 can be 45 ° (a crossing angle of 40 ° around 25 °). By setting the contact angle in this way, the radial load and the axial load can be received in a well-balanced manner, so that high rigidity can be exhibited with respect to the load in either direction.

  The second bearing 300 is a single-row angular contact ball bearing. The second bearing 300 may be a two-point contact, but is preferably a multi-point contact (four-point contact in the present embodiment). The second bearing 300 shown in FIG. 2 is a four-point contact, and the balls 306 are disposed in a single row of raceways formed between the outer ring 302 and the inner ring 304.

  The contact angle of the second bearing 300 is preferably the same as that of the gear-side row of the first bearing 200. That is, the contact angles θ4 and θ5 between the ball 306 and the raceway surface are both set to one side of a line (broken line) perpendicular to the pinion shaft 106 (gear shaft). The contact angle θ4 of the ball 306 is in the range of 5 ° to 15 °, and the contact angle θ5 is in the range of 40 ° to 50 °. As an example, the contact angle θ4 can be 5 °, and the contact angle θ5 can be 45 ° (the crossing angle is 40 ° around 25 °).

  Next, specific description will be made using examples and comparative examples. FIG. 3 is a view for explaining the configuration of the embodiment, FIG. 4 is a view for explaining the configuration of the comparative example, and FIG. 5 is a diagram for comparing the embodiment with the comparative example. About the part which description overlaps with FIG. 1, FIG. 2, the same code is attached and description is abbreviate | omitted.

  In Example 1 shown in FIG. 3A, the first bearing 200 is a double-row angular contact ball bearing. The gear row of the double row is in four-point contact, and the contact angles are 5 ° and 45 °. The other row of the double row was a two point contact and the contact angle was 40 °. The second bearing 300 is a single-row angular contact ball bearing. The second bearing 300 was in four-point contact, and the contact angles were 5 ° and 45 °.

  The second embodiment shown in FIG. 3B differs from the first embodiment in that the second bearing 300a is in two-point contact. The contact angle of the second bearing 300a was 40 °. The first bearing 200 has the same configuration as that of the first embodiment.

  The third embodiment shown in FIG. 3C is different from the second embodiment in that the gear row of the double row of the first bearings 200 is in two-point contact and the flange side is in four-point contact. There is. The contact angle on the gear side was 35 °, and the contact angles on the flange side were 5 ° and 45 °. The second bearing 300 a has the same configuration as that of the second embodiment.

  In Comparative Example 1 shown in FIG. 4 (a), the first bearing 350a and the second bearing 350b are both tapered roller bearings. The first bearing 350 a is disposed in the vicinity of the pinion gear 104 with the front facing the gear side. The second bearing 350 b is disposed in the vicinity of the flange 108 (on the side of the pinion shaft 106 opposite to the pinion gear 104) with the front facing the flange 108 side.

  Although the 1st bearing 352a is a double row angular contact ball bearing in the comparative example 2 shown in FIG.4 (b), both rows are 2 point contact. The first bearing 352 a has a contact angle of 40 ° in both of the double rows. The second bearing 352 b is a single-row angular contact ball bearing having a two-point contact, and has a contact angle of 24 °.

  In Comparative Example 3 shown in FIG. 4C, the first bearing 354a has the same configuration as the first bearing 352a of Comparative Example 2. The second bearing 352 b is a two-point contact single-row angular contact ball bearing, and the contact angle is 40 ° (the same configuration as in the second embodiment).

  In Comparative Example 4 shown in FIG. 4D, the first bearing 356a is a double row angular contact ball bearing, and both rows of the double row are in four point contact, and the contact angles are respectively 5 ° and 45 °. The second bearing 356 b is a two-point contact single-row angular contact ball bearing, and the contact angle is 40 ° (the same configuration as in the second embodiment).

  FIG. 5 is a diagram comparing loss and rigidity in the example and the comparative example. FIG. 5A shows the loss torque ratio based on the tapered roller bearing of Comparative Example 1. FIG. 5B shows the gear meshing position displacement amount ratio and the connection position displacement amount ratio based on the tapered roller bearing of Comparative Example 1 as the rigidity.

  Referring to the loss torque ratio of FIG. 5A, it can be seen that the loss torque ratio is smaller in Examples 1-3 and Comparative Examples 2-4, which are ball bearings, as compared to Comparative Example 1 of the tapered roller bearing. . Among them, in Comparative Examples 2 and 3 in which the first bearings in the double row do not have multipoint contact, the loss torque ratio is significantly small. On the other hand, in Comparative Example 4 in which the first bearing is in multi-point contact in both rows, it can be seen that the loss torque ratio is the largest among the ball bearings.

  Referring to the gear meshing position displacement amount ratio shown in FIG. 5 (b), in Comparative Examples 2 and 3 in which the double-row first bearing is not multipoint contact, the displacement amount ratio is significantly large and the rigidity is insufficient. I understand that. In the first and second embodiments in which the gear side of the double row is in multipoint contact, the gear mesh position displacement amount ratio is substantially the same as that of the first comparative example. In Example 3 in which the flange side of the double row is in multi-point contact, a slight decrease in rigidity can be seen, although not as in Comparative Examples 2 and 3. In Comparative Example 4 in which both double rows are in multipoint contact, the rigidity is higher than that in Comparative Example 1.

  From these facts, it is understood that although the loss torque is reduced when replacing the tapered roller bearings with the angular ball bearings, the rigidity is insufficient only by double rows as in Comparative Examples 2 and 3. Therefore, with the configurations of Comparative Examples 2 and 3, it is necessary to increase the size of the first bearing in order to secure rigidity. On the other hand, if double-row double-row rows are made to be multipoint contact as in Comparative Example 4, although rigidity can be secured, reduction of loss torque is small and there is little merit to replace tapered roller bearings.

  On the other hand, as in Example 1-3, the first bearing 200 on the gear side is a double-row angular contact ball bearing, and one row of the double rows is multipoint contact, thereby reducing loss torque and lowering rigidity. It can be seen that it is possible to achieve both suppression of Therefore, with the configuration of the embodiment 1-3, the fuel efficiency of the vehicle can be improved, and the size of the parts is not increased, so that the freedom of design can be secured.

  Incidentally, when the second and third embodiments are compared, the gear meshing position displacement ratio is smaller in the second embodiment in which the gear side is multipoint contact than in the third embodiment in which the flange side is multipoint contact. Therefore, it is understood that the configurations of the first and second embodiments are more preferable than the configuration of the third embodiment.

  In the connection position displacement amount ratio shown in FIG. 5B, the influence of the configuration of the second bearing is dominant. In Examples 2 and 3 and Comparative Examples 3 and 4 in which the second bearing is in two-point contact and the contact angle is 40 °, it can be seen that the displacement ratio is significantly large and the rigidity is insufficient. Although the second bearing 352b of Comparative Example 2 is a two-point contact, it is considered that the rigidity against a radial load is high because the contact angle is as small as 24 °. However, in the configuration of Comparative Example 2, it is apparent that the rigidity against axial load is reduced. On the other hand, in Example 1 in which the second bearing is in four-point contact, the connection position displacement amount ratio can be suppressed to be almost as small as Comparative Example 1. Therefore, if it is the structure of Example 1, since it is possible to replace also in a 2nd bearing, without size-up, it turns out that it is more preferable than the structure of Examples 2 and 3.

  In the above-mentioned embodiment, a pinion shaft was mentioned as an example and explained as a target which a bearing mechanism supports. However, the present invention is not limited to this, and the present invention can be applied to any bearing mechanism that supports a gear shaft having a gear at one end, and the benefits of the present invention can be enjoyed.

  In addition, although the example in which the first bearing 200 is a four-point contact as a multipoint contact has been described, it may be a three-point contact.

  Although the preferred embodiments of the present invention have been described above with reference to the accompanying drawings, it goes without saying that the present invention is not limited to such examples. It will be apparent to those skilled in the art that various changes or modifications can be conceived within the scope of the appended claims, and of course these also fall within the technical scope of the present invention. It is understood.

  The present invention can be used as a bearing mechanism for supporting a gear shaft having a gear at one end, such as a pinion shaft of a differential gear of a motor vehicle, for example.

102: ring gear, 104: pinion gear, 106: pinion shaft, 108: flange, 200: first bearing, 202: outer ring, 204: inner ring, 206a: ball, 206b: ball, 300: second bearing, 302: outer ring, 304 ... inner ring, 306 ... ball, 350a, 352a, 354a, 356a ... first bearing, 300a, 350b, 352b, 356b ... second bearing

Claims (3)

A bearing mechanism for supporting a gear shaft provided with a gear at one end,
A first bearing disposed near the gear;
And a second bearing disposed on the opposite side of the gear shaft from the gear shaft,
A bearing mechanism characterized in that the first bearing is a double-row angular contact ball bearing and only a row on the gear side of the double-row is multipoint contact.   The bearing mechanism according to claim 1, wherein the second bearing is a multi-point contact single-row angular contact ball bearing.   The row of multipoint contacts of the first bearing is a four-point contact, and the contact angles θ1 and θ2 between the raceway surface and the ball are both on one side of a line perpendicular to the gear axis, and the contact angle θ1 is The bearing mechanism according to claim 1 or 2, wherein the contact angle θ2 is in the range of 5 ° to 15 ° and 40 ° to 50 °.

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