JP2016011673A - Bearing mechanism - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a bearing mechanism capable of suppressing rigidity degradation while reducing loss torque than conventional tapered roller bearings.SOLUTION: A representative configuration of the invention is a bearing mechanism to pivotally support a gear shaft having a gear at one end thereof. The bearing mechanism includes a first bearing 200 disposed in the vicinity of the gear, and a second bearing 300 disposed opposite the gear of the gear shaft. The first bearing 200 is a multirow angular ball bearing. One of the multirow is of multipoint contact.

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 an automobile.

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

  However, although the tapered roller bearing has an advantage of high rigidity, there is a problem that the rotational resistance is large and the loss torque is large. Therefore, Patent Document 1 proposes a configuration in which, among 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 stated that the above configuration can reduce the rotational torque and improve the fuel efficiency of the automobile.

Japanese Patent No. 4250952

  Angular ball bearings are known to be able to receive both radial and axial loads. However, even if the angular ball bearings are arranged in a double row, a reduction in rigidity is inevitable as compared with a tapered roller bearing. Since a very large load is applied to the pinion shaft, the size of the bearing is inevitably increased when trying to have the same rigidity as the tapered roller bearing. When the size of the bearing is increased, there are problems that the fuel consumption is reduced due to an increase in the weight of the vehicle, and that the degree of freedom in component layout is reduced in designing the vehicle.

  Accordingly, an object of the present invention is to propose a bearing mechanism capable of suppressing a reduction in rigidity while reducing a loss torque as compared with a conventional tapered roller bearing.

  In order to solve the above problems, a representative configuration of the present invention is a bearing mechanism that supports a gear shaft having a gear at one end, and includes a first bearing disposed in the vicinity of the gear and a gear shaft. And a second bearing disposed on the opposite side of the gear, wherein the first bearing is a double-row angular ball bearing, and one of the double rows is in a multipoint contact.

  According to the above configuration, since it is a ball bearing, it is possible to suppress a decrease in rigidity, so that the loss torque can be reduced without increasing the size from the tapered roller bearing and more than the tapered roller bearing. In addition, since the rigidity is high compared to the case of using a single row or double row angular contact ball bearing, the amount of gear meshing point displacement can be kept small. Multipoint contact means that the ball and the raceway are in contact at three or more points.

  The second bearing is preferably a multi-point contact single row angular contact ball bearing. Also in the second bearing, by making multipoint contact, the amount of displacement of the connecting portion (for example, companion flange) on the opposite side of the gear can be kept small.

  The multi-point contact row 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 shaft, and the contact angle θ1 is 5 °. To 15 °, and 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 according to the present invention, it is possible to suppress the loss of rigidity while reducing the loss torque as compared with the conventional tapered roller bearing.

It is a fragmentary sectional view explaining the bearing mechanism concerning an embodiment. It is a figure explaining the detailed structure of a bearing. It is a figure explaining the structure of an Example. It is a figure explaining the structure 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 shown in the embodiments are merely examples for facilitating understanding of the invention, and do not limit the present invention unless otherwise specified. In the present specification and drawings, elements having substantially the same functions and configurations are denoted by the same reference numerals, and redundant description is omitted, and elements not directly related to the present invention are illustrated or described. Is omitted.

  FIG. 1 is a partial cross-sectional view illustrating a bearing mechanism according to an embodiment. In FIG. 1, a ring gear 102 of a differential gear and a pinion gear 104 are engaged 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 a 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 the front facing the gear side. The second bearing 300 is disposed near 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 includes a first bearing 200 and a second bearing 300.

  FIG. 2 is a diagram illustrating a 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 of the double rows is in multipoint contact. In the example of FIG. 2, the gear-side row of the double rows of the first bearings 200 is a four-point contact.

  In the first bearing 200, balls 206 a and 206 b are arranged in a double row track formed between the outer ring 202 and the inner ring 204. The contact angles θ1 and θ2 between the balls 206a in the gear-side row and the raceway surface are both on one side of a line (indicated 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 preferably 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 ° (the crossing angle is 40 ° with 25 ° as the center). 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 exerted against the load in either direction.

  The second bearing 300 is a single row angular ball bearing. The second bearing 300 may be two-point contact, but is preferably multipoint contact (four-point contact in this embodiment). The second bearing 300 shown in FIG. 2 has a four-point contact, and the balls 306 are arranged in a single row track 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 first bearing 200 on the gear side. That is, the contact angles θ4 and θ5 between the ball 306 and the raceway surface are both set on one side of a line (broken line) perpendicular to the pinion shaft 106 (gear shaft). The contact angle θ4 of the ball 306 is 5 ° to 15 °, and the contact angle θ5 is 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 set to 40 ° with 25 ° as the center).

  Next, it demonstrates concretely using an Example and a comparative example. FIG. 3 is a diagram illustrating the configuration of the example, FIG. 4 is a diagram illustrating the configuration of the comparative example, and FIG. 5 is a diagram comparing the example and the comparative example. Parts that are the same as those in FIGS. 1 and 2 are denoted by the same reference numerals and description thereof is omitted.

  In Example 1 shown in FIG. 3A, the first bearing 200 is a double-row angular ball bearing. Of the double row, the row on the gear side is a 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 ball bearing. The second bearing 300 is a four-point contact, and the contact angles are 5 ° and 45 °.

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

  In Comparative Example 1 shown in FIG. 4A, both the first bearing 350a and the second bearing 350b are tapered roller bearings. The first bearing 350a is disposed in the vicinity of the pinion gear 104 with the front facing the gear side. The second bearing 350b 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.

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

  In the comparative example 4 shown in FIG.4 (d), the 1st bearing 356a was a double row angular contact ball bearing, both rows were made into 4 point contact among double rows, and the contact angles were 5 degrees and 45 degrees, respectively. The second bearing 356b is a two-point contact single-row angular contact ball bearing and has a contact angle of 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 a loss torque ratio based on the tapered roller bearing of Comparative Example 1. FIG. 5B shows the gear engagement position displacement ratio and the connection position displacement ratio based on the tapered roller bearing of Comparative Example 1 as rigidity.

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

  Referring to the gear meshing position displacement ratio shown in FIG. 5B, Comparative Examples 2 and 3 in which the first bearings in the double row are not in multipoint contact have a significantly large displacement ratio and lack of rigidity. I understand that. In Examples 1 and 2 in which the double row gear side is in multipoint contact, the gear meshing position displacement amount ratio is substantially the same as that in Comparative Example 1. In Example 3 in which the double-row flange side is in multipoint contact, although not as much as Comparative Examples 2 and 3, a slight decrease in rigidity is observed. In Comparative Example 4 in which both of the double rows are in multipoint contact, the rigidity is higher than in Comparative Example 1.

  From these facts, it can be seen that, when the tapered roller bearing is replaced with the angular ball bearing, the loss torque is reduced, but the rigidity is insufficient only by the double row as in Comparative Examples 2 and 3. For this reason, in the configurations of Comparative Examples 2 and 3, it is necessary to increase the size of the first bearing in order to ensure rigidity. On the other hand, when both rows in the double row are in multipoint contact as in Comparative Example 4, although rigidity can be ensured, there is little reduction in loss torque and there are few merits to replace the tapered roller bearing.

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

  When comparing the second and third embodiments, the gear engagement position displacement amount 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 can be seen that the configurations of Examples 1 and 2 are more preferable than the configuration of Example 3.

  In the connection position displacement amount ratio shown in FIG. 5B, the influence of the configuration of the second bearing is dominant. It can be seen that in Examples 2 and 3, and Comparative Examples 3 and 4 in which the second bearing is in a two-point contact and the contact angle is 40 °, the displacement amount ratio is significantly large and the rigidity is insufficient. Although the 2nd bearing 352b of the comparative example 2 is 2 point | piece contact, since the contact angle is as small as 24 degrees, it seems that the rigidity with respect to radial load was high. However, in the configuration of Comparative Example 2, it is clear that the rigidity against the axial load is lowered. On the other hand, in Example 1 in which the second bearing is in a four-point contact, the connection position displacement amount ratio can be kept almost as small as in Comparative Example 1. Therefore, it can be seen that the configuration of the first embodiment is more preferable than the configurations of the second and third embodiments because the second bearing can be replaced without increasing the size.

  In the above embodiment, the pinion shaft has been described as an example of the object supported by the bearing mechanism. 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 can benefit from the present invention.

  Moreover, although the example which made the 1st bearing 200 4 point contact as multipoint contact was demonstrated, it is good also as 3 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 and modifications can be made within the scope of the claims, and these are naturally within the technical scope of the present invention. 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 an automobile.

DESCRIPTION OF SYMBOLS 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 in the vicinity of the gear;
A second bearing disposed on the opposite side of the gear shaft from the gear;
The first bearing is a double row angular contact ball bearing, and one of the double rows is a multipoint contact.   The bearing mechanism according to claim 1, wherein the second bearing is a multi-contact single-row angular contact ball bearing.   The multi-point contact row 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 shaft, 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 the contact angle θ2 is 40 ° to 50 °.

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