JP3359501B2 - Tapered roller bearing for pinion shaft support of differential gear - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to an improvement in a tapered roller bearing for rotatably supporting a pinion shaft constituting a differential gear (final reduction gear) of an automobile inside a casing (differential case).

[0002]

2. Description of the Related Art A differential gear which is provided in the middle of a power transmission system of an automobile to reduce the rotation of a propeller shaft and convert the rotation direction into a right angle at the same time is configured as shown in FIG. Inside front of casing 1 (to the right in FIG. 1)
A pinion shaft 2 is provided in the portion. A rear end of a propeller shaft (not shown) can be connected to a coupling flange 3 fixed to a portion of the front end of the pinion shaft 2 (the right end in FIG. 1) protruding from the front end opening of the casing 1. Also, the rear end of the pinion shaft 2 (the left end in FIG. 1).
, A reduction gear 4 is fixed, and the reduction gear 4 and the reduction gear 5 mesh with each other. This reduction gear 5
Is rotatably supported inside the rear part (left part of FIG. 1) of the casing 1. Further, two positions in the front and rear of the intermediate portion of the pinion shaft 2 correspond to a pair of front and rear tapered roller bearings 6.
The casing 1 is rotatably supported by a and 6b.

Each of these tapered roller bearings 6a, 6b has one outer ring 7a, 7b and one inner ring 8a, 8b, respectively.
Each is composed of a plurality of tapered rollers 9a and 9b. Conical concave outer raceways 10a and 10b are formed on the inner peripheral surfaces of the outer races 7a and 7b, and conical convex inner raceways 11a and 11b are formed on the outer peripheral surfaces of the inner races 8a and 8b, respectively. The outer races 7a and 7b are internally fitted and fixed to a part of the casing 1, and the inner races 8a and 8b are externally fitted and fixed at two positions in front and rear of an intermediate portion of the pinion shaft 2.

FIG. 2 shows a tapered roller bearing 6a for supporting the pinion shaft 2 (FIG. 1) of such a differential gear.
(The same applies to 6b). Tapered roller bearing 6a incorporated in a conventional differential gear
Is the outer ring raceway 10a is longitudinal contact angle α is 20 degrees is an angle which is inclined with respect to the center axis of the outer ring 7a, the length of the diameter D _a and the tapered rollers 9a of the large-diameter end of the tapered rollers 9a is the ratio D _a / L and L is 0.4 to 0.9, the number coefficient k time was about 1.1 to 1.3. The inner diameter of the inner ring 8a differs depending on the type of automobile and before and after the mounting position. In the case of the tapered roller bearing to which the present invention is applied, the inner diameter is about 55 mm or less.

[0005] The roller number coefficient k is defined as the outer ring raceway 10.
a (10b) and the center of the conical rollers 9a (9b) arranged at equal intervals on the pitch circle to indicate the degree of consolidation of the conical rollers 9a (9b) between the inner ring raceway 11a (11b). it is the inverse of the ratio of the diameter D _a of the larger diameter end of the tapered rollers 9a (9b) with respect to the distance, is represented by the following formula. k = (d _m / D _a ) · sin (180 ° / z) where d _m is the diameter (at the large-diameter side end) of the pitch circle of the plurality of tapered rollers 9a (9b). z is tapered roller 9a
(9b) represents the number. k = 1 indicates a state in which the tapered rollers 9a (9b) are tightly packed,
As the value of k increases, the number of tapered rollers 9a (9b) decreases.

[0006]

In the case of the tapered roller bearing for supporting the pinion shaft of the conventional differential gear configured as described above, the tapered roller bearing incorporated in the differential gear is required in terms of fatigue life, bearing rigidity, and the like. Although the minimum performance was sufficiently satisfied, the rotational torque was not always small enough. In recent years, there has been an increasing demand for fuel efficiency in automobiles, and in the case of the tapered roller bearings as well, it has been required to reduce the rotational torque in order to suppress the power transmission loss. However, it is not preferable that the fatigue life and the bearing rigidity are excessively reduced by reducing the rotational torque. In view of such circumstances, the present invention has been conceived to obtain a tapered roller bearing for supporting a pinion shaft of a differential gear having a small rotational torque while ensuring fatigue life and bearing rigidity.

[0007]

Processes leading to the present invention The following four factors can be considered as factors affecting the fatigue life, bearing stiffness, and rotational torque of tapered roller bearings. Roller number coefficient k diameter of the tapered rollers of the large diameter end D _a of the tapered rollers of the large diameter side end portion diameter D _a and the ratio between the length L of the tapered roller D _a / L contact angle α

Therefore, the effects of these elements on the fatigue life, bearing rigidity (axial rigidity and radial rigidity), and rotational torque (low torque property) are considered, and the results are shown in Table 1 below. Indicated. The arrows in Table 1 indicate the effect of the element on the performance.
The upward arrow indicates that the performance is improved (the fatigue life is long, the bearing rigidity is high, the rotational torque is low, and the low torque property is improved) as the value of the element increases.
Conversely, a downward arrow indicates that the performance decreases as the value of the element increases (short fatigue life, low bearing stiffness, high rotational torque, and low torque performance). Further, the numerical values described in the rightmost column in this table show a range of values that can be adopted as the tapered roller bearing for supporting the pinion shaft of the differential gear. Roller number coefficient k
And the ratio D _a / L is a unit of dimensionless diameter D _a is mm, the contact angle unit of α in degrees (°). An axial load and a radial load are applied to a tapered roller bearing for a differential gear during use. The following Table 1 was prepared assuming that the magnitudes of the loads in these two directions were the same.

[0009]

[Table 1]

As is clear from the description in Table 1, by simply changing the roller number coefficient k, the ratio D _a / L, the diameter D _a , and the contact angle α, the fatigue life, the bearing rigidity, and the
It is not possible to improve the performance related to the rotational torque at the same time. Therefore, the inventor of the present invention has conducted many experiments and invented a tapered roller bearing for supporting a pinion shaft of a differential gear capable of securing a practically sufficient fatigue life and bearing stiffness and reducing rotation torque.

[0011]

SUMMARY OF THE INVENTION A tapered roller bearing for supporting a pinion shaft of a differential gear according to the present invention has a front end formed on a propeller shaft in the same manner as a conventionally known tapered roller bearing for supporting a pinion shaft of a differential gear. The pinion shaft is rotatable with respect to the casing at two positions before and after an intermediate portion of a pinion shaft having a rear end portion that is freely connectable and a reduction small gear meshing with the reduction gear wheel fixed to the rear end portion.

Particularly, in the tapered roller bearing for supporting the pinion shaft of the differential gear according to the present invention, the contact angle α is 2
Is 2 to 28 degrees, the ratio D _a / L of the diameter D _a and the tapered rollers of the length L of the larger diameter end of the tapered rollers is from 0.51 to 1.0, the pitch of the plurality of tapered rollers the diameter of a circle and d _m, the number of tapered rollers when the _{z, k = (d m /} D a) · si
The roller number coefficient k represented by n (180 ° / z) is 1.16
~ 1.32.

[0013]

In the case of the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention configured as described above, the rotational torque can be reduced while ensuring a practically sufficient fatigue life and bearing rigidity. Therefore, while ensuring the performance required for the differential gear, the power loss at the differential gear portion can be reduced, thereby contributing to fuel saving of the automobile.

If the contact angle α is controlled in the range of 26 to 28 degrees and the ratio D _a / L is controlled in the range of 0.7 to 1.0, the effect of reducing the rotational torque is further improved.
Also, if regulated when regulating the contact angle α in the range of 22 to 24 degrees at the same time the ratio D _a / L in the range of 0.51 to 0.8, when regulating the contact angle α in the range of 26 to 28 degrees Although the effect of reducing the rotational torque is smaller than that of, the radial rigidity can be sufficiently secured. If the contact angle α is restricted to the range of 24 to 26 degrees, an intermediate characteristic can be obtained. Further, the surface roughness of the large-diameter end surface of each tapered roller and the surface on which the large-diameter end surface slides on one surface of the flange formed on the large-diameter outer peripheral surface of the inner ring are each 0.1 μRa or less. If not restricted, the rotational torque can be further reduced and, at the same time, the seizure resistance can be improved.

[0015]

EXAMPLES In order to confirm the effects of the present invention, a part of experiments performed by the present inventors will be described. Tables 2 and 3 below
Shows the specifications of the six types of tapered roller bearings used in this experiment. Of these, three types of tapered roller bearings A to C are located at the front of the intermediate portion of the pinion shaft 2 (FIG. 1) (right side in FIG. 1).
, And three types of tapered roller bearings D to F, which also support the rear side of the middle part (the left side in FIG. 1), have relatively large diameters. . or,
The two types of tapered roller bearings A and D are different from those conventionally incorporated in differential gears in B, C, E and F.
The four types of tapered roller bearings described above belong to the present invention. In the tapered roller bearing used in the experiment, SUJ 2 (JIS G 4805), which is a high carbon chromium bearing steel, was used for each of the outer ring, the inner ring, and the tapered rollers. However, these members are made of chrome steel SC
r 420H, SCr 430H, SCr 440H
(JIS G 4052), medium carbon molybdenum steel, medium carbon chromium manganese steel. Further, the materials of the constituent members need not always be the same, and different materials can be used in combination for each member.

[0016]

[Table 2]

[Table 3]

[0017] Studies on the number of tapered roller bearings including six tapered roller bearing shown in these Tables 2 and 3, using the experimental apparatus as shown in FIG. 3, fatigue life, bearing stiffness, The rotation torque was measured. The holder 12 to which the inner race 8a (or 8b) of the tapered roller bearing 6a (or 6b) is externally fitted and fixed is taperedly fitted to the upper end of the drive shaft 13 and is driven to rotate by the drive shaft 13. The outer ring 7a (or 7b) is internally fitted and fixed inside the housing 15 via the outer holder 14. A predetermined lubricating oil can be supplied into the housing 15 through an oil supply hole 16. A predetermined axial load can be applied to the upper surface of the housing 15 via a static pressure pad 17. Further, a load cell 19 is provided between the distal end of the arm piece 18 fixed to the outer peripheral surface of the housing 15 and a fixed portion (not shown), so that the dynamic torque (==) applied to the housing 15 when the drive shaft 13 rotates. The rotation torque of the tapered roller bearing 6a (or 6b) can be measured freely. The purpose of reducing the dynamic torque is to achieve fuel saving as described above. Therefore, the radial load, which has little effect on fuel economy,
It was not applied when measuring dynamic torque. That is, during operation of the differential gear, the thrust load is always applied, but a large radial load is applied only in limited cases such as sudden acceleration / deceleration, and the proportion of the total operation time is small. Therefore, the influence of the change in dynamic torque due to the radial load on the fuel economy performance is smaller than that of the thrust load. Therefore, the load applied at the time of the dynamic torque measurement experiment is
Only the thrust load was used.

As a first experimental result, FIG. 4 shows the relationship between the product of the inner diameter d of the tapered roller bearing and the set width W (see FIG. 2) and the dynamic torque (= rotation torque) of the tapered roller bearing. Shown in FIG. 4 shows the product of the inner diameter d and the set width W on the horizontal axis.
The ordinate represents the dynamic torque of the tapered roller bearing when the drive shaft 13 (FIG. 3) is rotated at 3000 rpm. The symbols A to F circled correspond to the leftmost columns of Tables 2 and 3 . The smaller the product is, the smaller the tapered roller bearing is. As is apparent from the description of FIG. 4, the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention has a smaller dynamic torque than the conventional product when the size is the same.

FIG. 5 shows the results of an experiment conducted to determine the relationship between the rotation speed and the dynamic torque for each of the tapered roller bearings A to F. Experiment # 90 and oil temperature 8
Supply 0 ° C gear oil at a rate of 600 cc / min.
The test was performed while applying an axial load of 0 kgf. FIG. 5A shows the results of experiments on a relatively small diameter tapered roller bearing that supports the front side of the intermediate portion of the pinion shaft 2.
(B) shows the result of an experiment on a relatively large-diameter tapered roller bearing that also supports the rear side of the intermediate portion.
FIG. 5 also shows that the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention has a smaller dynamic torque than the conventional product.

[0020] As described above, the tapered roller bearing of the present invention, the contact angle alpha, the ratio D _a / L of the diameter D _a and the tapered rollers of the length L of the larger diameter end of the tapered rollers, rollers By limiting the number coefficient k, the dynamic torque can be reduced. In order to further reduce this dynamic torque and improve seizure resistance so that seizure does not occur even in the case of extreme lubrication failure, the large-diameter side end face of each tapered roller and the large diameter of the inner ring It is effective to regulate the surface roughness of one surface of the flange formed on the outer peripheral surface to the surface on which the large-diameter side end surface slides, to 0.1 μRa or less. FIGS. 6 to 6 show the results of experiments performed to determine the effect of regulating the surface roughness of both surfaces.
7 will be described.

First, in an experiment on the effect of the surface roughness on dynamic torque, in Tables 2 and 3 , a tapered roller bearing having specifications corresponding to B in Tables 2 and 3 was used. Supply gear oil with a temperature of 80 ° C at a rate of 600 cc / min,
The test was performed while applying an axial load of 900 kgf. First, in FIG. 6 showing the relationship between the rotation speed and the dynamic torque, the broken line A shows the experimental result of the comparative example in which the surface roughness of both surfaces was 0.3 μRa, and the solid line B shows the surface roughness of both surfaces. To 0.
The experimental results for the case of 1 μRa are shown. As is clear from FIG. 6, by regulating the surface roughness of both surfaces to 0.1 μRa or less, the dynamic torque in the practical rotation range can be greatly reduced. In addition, FIG.
In the experiment whose results are shown in FIG. 5, the roughness on both surfaces was regulated to 0.1 μRa or less.

FIG. 7 shows the effect of the surface roughness of the two surfaces on the seizure resistance. Experiments on anti-seizure properties were carried out on tapered roller bearings to be tested with # 9.
0.2cc of 0 gear oil is applied, and it is rotated at 4000rpm while applying an axial load of 1000kgf.
The measurement was performed by measuring the time until image sticking occurred. In FIG. 7 showing the results of this experiment, the vertical axis represents the time (sec) until the occurrence of image sticking, and the horizontal axis represents the root-mean-square roughness of both surfaces [= ｛(surface roughness of the large-diameter end surface of the tapered roller). ² + (Surface roughness of the surface on one side of the flange where the large-diameter end surface slides) ² ｝
^1/2 ], respectively. Of the plurality of points described in FIG. 7, three triangles indicate the results of experiments on the tapered roller bearings in which the surface roughness on both sides was regulated to 0.1 μRa or less, and the remaining circles indicate the results on the both sides. The results of experiments on tapered roller bearings having a coarser surface roughness are shown. As is clear from the description of FIGS. 6 and 7, by regulating the surface roughness of both surfaces to 0.1 μRa or less, the dynamic torque in the practical rotation range is greatly reduced and the seizure resistance is improved. Can do things.

As described above, the tapered roller bearing for supporting the pinion shaft of the differential gear according to the present invention can reduce the dynamic torque as compared with the conventionally known tapered roller bearing for supporting the pinion shaft of the differential gear. However, practically sufficient values can be ensured for fatigue life and bearing rigidity. The fatigue life and the bearing rigidity can be obtained by calculation if the bearing data as described in Tables 2 and 3 are known.

FIG. 8 is a graph showing the values of A shown in Tables 2 and 3 above.
The results obtained by calculating the fatigue life of the six types of tapered roller bearings No. to No. F are shown. The vertical axis indicates the fatigue life of the conventional A and D double tapered roller bearings as 1, and the life of the other tapered roller bearings as a percentage of the life of these A and D double tapered roller bearings. As is apparent from FIG. 8, the tapered roller bearing for supporting the pinion shaft of the differential gear according to the present invention can secure the same fatigue life as the conventional product.

Next, FIG. 9 shows the six types of tapered roller bearings A to F described in Tables 2 and 3 above, in which a combination of A and D, a combination of B and E, and a combination of C and F The combination shows the axial bearing stiffness in a state where the pinion shaft 2 (FIG. 1) is mounted at two positions before and after the intermediate portion and the pinion shaft 2 is rotatably supported. The vertical axis represents the axial bearing stiffness (displacement) when the pinion shaft 2 is supported by the conventional A and D double tapered roller bearings, and the axial bearing stiffness when the other tapered roller bearings are combined. It is described as a ratio to the axial bearing stiffness when the double tapered roller bearings A and D are combined. Since the value on the vertical axis is expressed as a ratio of the amount of displacement when a predetermined axial load is applied, the lower the numerical value, the higher the bearing rigidity. As is apparent from FIG. 9, when the pinion shaft 2 is supported by the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention, an axial bearing rigidity equal to or higher than that of the conventional case can be obtained.

Next, FIG. 10 shows the result of calculating the radial bearing stiffness under the same conditions as in FIG. FIG. 9 above
Similarly to the above case, the value on the vertical axis is expressed as a ratio of the amount of displacement when a predetermined radial load is applied, so that the lower the numerical value, the higher the bearing rigidity. As apparent from FIG. 10, when the pinion shaft 2 is supported by the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention,
Radial bearing rigidity is lower than in the conventional case. Therefore, when the B and E double tapered roller bearings are combined, as in the case of combining the conventional A and D double tapered roller bearings, the present invention can be applied to a differential gear of a vehicle equipped with a high-power engine. When a tapered roller bearing is combined, the present invention can be applied to a differential gear of an automobile equipped with a low-output engine.

By considering the results of many of the experiments described above, including each of the experiments described above, the inventor has determined that the contact angle may be 2
In the range of 2 to 28 degrees, the ratio D _a / L of the diameter D _a of the larger diameter end of the tapered rollers which constitute the tapered roller bearing and the length L of the tapered roller of 0.51 to 1.0 A tapered roller bearing having a specific structure in which the roller number coefficient k is restricted to a range of 1.16 to 1.32 is particularly useful as a tapered roller bearing for supporting a pinion shaft of a differential gear. We came to think that action and effect could be obtained. Then, as shown in G1 to G8 in Table 4 below, with respect to the above three types of limiting elements (above), eight types of samples obtained by combining the above-described upper limit and lower limit, respectively, were used. The required performance of the tapered roller bearing for supporting the pinion shaft was measured, and it was found that sufficient performance could be obtained.

[0028]

[Table 4]

From the above description and the description in Table 4 , the scope of the present invention is a parallelepiped (for example, a rectangular parallelepiped or a cube) whose eight vertices are represented by G1 to G8 in Table 4. You can see that it is represented inside. In addition, these G1 to G8
The surface roughness between the large-diameter end face of each tapered roller and the surface on which the large-diameter end face slides on one side of the flange formed on the large-diameter outer peripheral face of the inner ring for the eight types of samples specified in. It was also found that good results could be obtained, as in the case of B, C, E, and F, if all were regulated to 0.1 μRa or less.

[0030]

As described above, the tapered roller bearing for supporting the pinion shaft of the differential gear according to the present invention is constructed and operates as described above. Therefore, while maintaining the performance required for the differential gear, the power at the differential gear portion is ensured. Loss can be reduced, which contributes to fuel efficiency of automobiles.

[Brief description of the drawings]

FIG. 1 is a vertical sectional side view showing an example of a differential gear.

FIG. 2 is a half sectional view showing an example of a tapered roller bearing.

FIG. 3 is a longitudinal sectional view of the experimental apparatus.

FIG. 4 is a diagram showing the relationship between the size of a tapered roller bearing and dynamic torque.

FIG. 5 is a diagram showing a relationship between a rotation speed and a dynamic torque.

FIG. 6 is a diagram showing the relationship between the rotational speed and the dynamic torque, showing the effect of the surface roughness of the sliding contact surface on the dynamic torque.

FIG. 7 is a diagram showing the effect of the surface roughness of the sliding contact surface on seizure resistance.

FIG. 8 is a graph showing the ratio of fatigue life.

FIG. 9 is a graph showing the ratio of axial bearing stiffness.

FIG. 10 is a graph showing a ratio of radial bearing rigidity.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Casing 2 Pinion shaft 3 Connection flange 4 Reduction gear 5 Reduction large gear 6a, 6b Tapered roller bearing 7a, 7b Outer ring 8a, 8b Inner ring 9a, 9b Tapered roller 10a, 10b Outer ring track 11a, 11b Inner ring track 12 Holder 13 Drive shaft 14 outer holder 15 housing 16 oil supply hole 17 static pressure pad 18 arm piece 19 load cell

──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Akira Ishimaru 1-5-50 Kumeinuma Shinmei, Fujisawa-shi, Kanagawa Nippon Seiko Co., Ltd. (56) References JP-A 1-156353 (JP, U) JP-A 58-58115 (JP, U) Actually open Showa 49-130247 (JP, U) (58) Fields investigated (Int. Cl. ⁷ , DB name) F16H 57/02 302 F16C 19/34

Claims (4)

(57) [Claims] 1. A front end part is freely connectable to a rear end part of a propeller shaft, and a rear end part is fixed with a reduction gear that meshes with a reduction gear. in supporting the pinion shaft for a tapered roller bearing of differential gear which rotatably supports a contact angle α is 22 to 28 degrees, the diameter D _a of the larger diameter end of the tapered rollers
And the ratio D _a / L of the tapered roller to the length L is 0.51 to 1.0
, And the diameter of the pitch circle of the plurality of tapered rollers and d _m,
When the number of tapered rollers is z, k = (d _m / D _a ) ·
The roller number coefficient k represented by sin (180 ° / z) is 1.1
6. A tapered roller bearing for supporting a pinion shaft of a differential gear, wherein the diameter of the tapered roller bearing is 6 to 1.32. 2. The contact angle α is 26 to 28 degrees and the ratio D _a
The tapered roller bearing for supporting a pinion shaft of a differential gear according to claim 1, wherein / L is 0.7 to 1.0. 3. The contact angle α is 22 to 24 degrees and the ratio D _a
The tapered roller bearing for supporting a pinion shaft of a differential gear according to claim 1, wherein / L is 0.51 to 0.8. 4. The surface roughness of the large-diameter end surface of each tapered roller and the surface of the flange portion formed on the large-diameter outer peripheral surface of the inner ring on which the large-diameter end surface slides are all zero. The tapered roller bearing for supporting a pinion shaft of a differential gear according to any one of claims 1 to 3, wherein the diameter of the tapered roller bearing is 0.1 μRa or less.