JPH0996352A - Pinion shaft supporting tapered roller bearing for differential gear - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce torque while both a practically sufficient fatigue life and bearing stiffness are being secured by making a contact angle specific, for the pinion shaft supporting tapered roller bearing for a differnetial gear. SOLUTION: In the pinion shaft supporting tapered roller bearing, a contact angle α is 22 to 28 deg., a ratio Da/L of the diameter Da of the large diameter end part of a tapered roller to the length of the tapered roller is 0.51 to 1.0, and furthermore, a roller number factor k represented by k=(dm/Da)×sin (180 deg.C/Z) is 1.16 to 1.32 where the diameter of a pitch circle for a plurality of tapered rollers is dm, and the number of the tapered rollers are Z. Still more, if a contact angle α is limited to a range of 26 deg. to 28 deg., and if a ratio of Da/L is simultaneously limited to a range of 0.7 to 1.0, a reduction effect of torque can thereby be enhanced much more. In addition, if the contact angle αis limited to a range of 22 to 24 deg., and if the ratio of Da/L is simultaneously limited to a range of 0.5 to 0.8, the reduction effect of torque is smaller as compared with a case where the contact angle α is limited to a range of 26 to 28 deg., however, its radial stiffness can thereby be sufficiently secured.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an improvement in a tapered roller bearing for rotatably supporting a pinion shaft, which constitutes 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, reduces the rotation of a propeller shaft and at the same time converts the rotation direction into a right angle, is constructed as shown in FIG. Frontward of the inside of the casing 1 (rightward in FIG. 1)
A pinion shaft 2 is arranged in the portion. A rear end portion of a propeller shaft (not shown) can be connected to a joint flange 3 fixed to a portion of the front end portion (right end portion in FIG. 1) of the pinion shaft 2 protruding from the front end opening portion of the casing 1. Also, the rear end of the pinion shaft 2 (the left end in FIG. 1)
The reduction small gear 4 is fixed to the gear, and the reduction small gear 4 and the reduction large gear 5 are meshed with each other. This reduction gear 5
Is rotatably supported inside the rear part (left part in FIG. 1) of the casing 1. Further, two positions in the front and rear of the intermediate portion of the pinion shaft 2 are located in 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 and 6b has one outer ring 7a and 7b and one inner ring 8a and 8b, respectively.
Each of them is composed of a plurality of tapered rollers 9a and 9b. Conical concave outer ring raceways 10a, 10b are formed on the inner peripheral surfaces of the outer rings 7a, 7b, and conical convex inner ring raceways 11a, 11b are formed on the outer peripheral surfaces of the inner rings 8a, 8b, respectively. The outer rings 7a and 7b are fitted and fixed to a part of the casing 1, and the inner rings 8a and 8b are fitted and fixed to the pinion shaft 2 at two positions before and after the middle portion.

FIG. 2 shows a tapered roller bearing 6a for supporting the pinion shaft 2 (FIG. 1) of such a differential gear.
(6b is also the same) is taken out and shown. Tapered roller bearing 6a incorporated in a conventional differential gear
Indicates that the contact angle α, which is the angle at which the outer ring raceway 10a is inclined with respect to the central axis of the outer ring 7a, is about 20 degrees, the diameter D _a of the large diameter side end of the tapered roller 9a and the length of the tapered roller 9a. The ratio D _a / L with the length L was 0.4 to 0.9, and the roller number coefficient k was about 1.1 to 1.3. The inner diameter of the inner ring 8a varies depending on the type of vehicle and before and after the mounting position, but in the case of the tapered roller bearing which is the object of the present invention, it is about 55 mm or less.

The coefficient k of the number of rollers means the outer ring raceway 10
In order to show the degree to which the tapered rollers 9a (9b) are clogged between a (10b) and the inner ring raceway 11a (11b), the centers of the tapered rollers 9a (9b) arranged at equal intervals on the pitch circle are connected. It is the reciprocal of the ratio of the diameter D _a of the large diameter side end of the tapered roller 9a (9b) to the long distance, and is represented by the following equation. k = (d _m / D _a ) · sin (180 ° / z) In this equation, d _m is the diameter (at the end on the large diameter side) of the pitch circle of the plurality of tapered rollers 9a (9b), z is a tapered roller 9a
The numbers in (9b) are shown. k = 1 means that the tapered rollers 9a (9b) are closely packed,
The larger the value of k, the smaller the number of tapered rollers 9a (9b).

[0006]

In the case of the conventional tapered roller bearing for supporting the pinion shaft of the differential gear configured as described above, fatigue life, bearing rigidity, etc. are required for the tapered roller bearing incorporated in the differential gear. Although the minimum performance is sufficiently satisfied, the rotating torque was not always sufficiently small. In recent years, there has been a strong demand for reducing fuel consumption of automobiles, and the tapered roller bearing is also required to have a smaller rotational torque in order to keep the power transmission loss low. However, it is not preferable that the fatigue life and the bearing rigidity are excessively reduced by reducing the rotating torque. In view of such circumstances, the present invention has been devised in order 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 invention The following four factors can be considered as factors that affect the fatigue life, bearing rigidity, and rotational torque of tapered roller bearings. Roller number coefficient k Diameter D _a of large diameter side end of tapered roller D _a Ratio of diameter D _a of large diameter side end of tapered roller to length L of tapered roller D _a / L Contact angle α

Therefore, the influence of each of these elements on the fatigue life, bearing rigidity (axial rigidity and radial rigidity), and rotational torque (low torque) is considered, and the results are shown in Table 1 below. Indicated. The arrow in Table 1 indicates the influence of the element on the performance,
The upward arrow indicates that the performance is improved as the value of the element is increased (the fatigue life is long, the bearing rigidity is high, the rotational torque is low, and the low torque property is improved).
On the other hand, the downward arrow indicates that the performance decreases as the value of the element increases (the fatigue life is short, the bearing rigidity is low, the rotation torque is high, and the low torque property is low). Further, the numerical values shown in the rightmost column of this table indicate the 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 dimensionless, the unit of the diameter D _a is mm, and the unit of the contact angle α is degree (°). The tapered roller bearing for the differential gear is subjected to a load in the axial direction and a load in the radial direction during use. The following Table 1 was prepared assuming that the magnitudes of the loads in these two directions are the same.

[0009]

[Table 1]

As is clear from the description in Table 1, the fatigue life, bearing rigidity, and bearing rigidity of the tapered roller bearing can be obtained by simply changing the roller number coefficient k, the ratio D _a / L, the diameter D _a , and the contact angle α.
Performance related to rotational torque cannot be improved at the same time. Therefore, the present inventor has conducted many experiments and invented a tapered roller bearing for supporting a pinion shaft of a differential gear, which can secure practically sufficient fatigue life and bearing rigidity, and can reduce the rotating torque.

[0011]

A tapered roller bearing for supporting a pinion shaft of a differential gear according to the present invention has a front end portion of a propeller shaft similar to that of a tapered roller bearing for supporting a pinion shaft of a differential gear which has been conventionally known. A pinion shaft is rotatably supported at two positions in the front and rear of an intermediate portion of a pinion shaft, which is connectable to a rear end portion and has a small reduction gear that meshes with a large reduction gear.

Particularly, in the tapered roller bearing for supporting the pinion shaft of the differential gear of 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 When the diameter of the circle is d _m and the number of tapered rollers is 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 constructed as described above, the rotational torque can be reduced while ensuring the fatigue life and the bearing rigidity which are practically sufficient. Therefore, while ensuring the performance required for the differential gear, it is possible to reduce the power loss in this differential gear portion and contribute to the fuel saving of the vehicle.

If the contact angle α is regulated within the range of 26 to 28 degrees and the ratio D _a / L is regulated within the range of 0.7 to 1.0, the effect of reducing the rotational torque is further improved.
When the contact angle α is restricted to the range of 22 to 24 degrees and the ratio D _a / L is restricted to the range of 0.51 to 0.8, the contact angle α is restricted to the range of 26 to 28 degrees. Although the effect of reducing the rotating torque is smaller than that of, the radial rigidity can be sufficiently secured. If the contact angle α is regulated within the range of 24 to 26 degrees, intermediate characteristics can be obtained. Further, the surface roughness between the large diameter side end surface of each tapered roller and the surface of the flange portion formed on the large diameter side outer peripheral surface of the inner ring with which the large diameter side end surface is in sliding contact is 0.1 μRa or less. If it is regulated to, 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 effect of the present invention, a part of the experiment conducted by the present inventor will be described. Table 2 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 those of a relatively small diameter that support the front side of the intermediate portion (right side of FIG. 1) of the pinion shaft 2 (FIG. 1) and three types of D to F. The tapered roller bearings each have a relatively large diameter and also support the rear side of the intermediate portion (left side in FIG. 1). Further, two kinds of tapered roller bearings A and D have been conventionally incorporated in a differential gear, and four kinds of tapered roller bearings B, C, E and F belong to the present invention. Shows. The tapered roller bearing used in the experiment was
SUJ 2 (JIS G 4805), which is a high carbon chromium bearing steel, was used for both the inner ring and the tapered rollers.
However, each of these members is made of chromium steel SCr 420.
H, SCr 430H, SCr 440H (JIS G
4052), medium carbon molybdenum steel, medium carbon chromium manganese steel. Further, the materials of the constituent members are not necessarily the same, and different materials may be used in combination for each member.

[0016]

[Table 2]

With respect to a large number of tapered roller bearings including the six types of tapered roller bearings shown in Table 2, fatigue life, bearing rigidity and rotational torque were measured by using an experimental device as shown in FIG. . The holder 12 to which the inner ring 8a (or 8b) of the tapered roller bearing 6a (or 6b) is externally fitted and fixed is taper-fitted to the upper end portion of the drive shaft 13 and is rotationally driven by the drive shaft 13. The outer ring 7 a (or 7 b) is fitted and fixed inside the housing 15 via the outer holder 14. A predetermined lubricating oil can be supplied into the housing 15 through the oil supply hole 16. Further, 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 a tip portion of the arm piece 18 fixed to the outer peripheral surface of the housing 15 and a fixed portion (not shown), and the drive shaft 1
The dynamic torque (=
The rotational torque of the tapered roller bearing 6a (or 6b) can be measured freely. The purpose of reducing the dynamic torque is to save fuel consumption as described above. Therefore, the radial load, which has little influence from the viewpoint of fuel saving, was not applied during the dynamic torque measurement. That is, when the differential gear is in operation, the thrust load is always applied, but a large radial load is applied in a limited case such as during sudden acceleration / deceleration, and the ratio to the entire operation time is small. Therefore, the influence of the dynamic torque change due to the radial load on the fuel economy performance is smaller than the thrust load. Therefore, the load applied during the dynamic torque measurement experiment was only the thrust load.

As a first experimental result, first, the relationship between the product of the inner diameter d of the tapered roller bearing and the assembly width W (see FIG. 2) and the dynamic torque (= rotation torque) of this tapered roller bearing is shown in FIG. Shown in. In FIG. 4, the horizontal axis is the product of the inner diameter d and the assembly width W,
The vertical axis represents the dynamic torque of the tapered roller bearing when the drive shaft 13 (FIG. 3) is rotated at 3000 rpm. Further, the symbols A to F surrounded by circles correspond to the leftmost column in Table 2 above. The smaller the product, the smaller the tapered roller bearing. As is clear from the description of FIG. 4, the dynamic torque of the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention is smaller than that of the conventional product when the sizes are the same.

Next, FIG. 5 shows the result of an experiment conducted for each of the tapered roller bearings A to F to obtain the relationship between the rotational speed and the dynamic torque. The experiment was # 90 and the oil temperature was 8
Supply gear oil at 0 ℃ at a rate of 600cc / min, and
It was performed while applying an axial load of 0 kgf. In FIG. 5, (A) shows the experimental results of a relatively small diameter tapered roller bearing that supports the front side of the intermediate portion of the pinion shaft 2.
(B) shows the experimental results of a relatively large diameter tapered roller bearing which also supports the rear side of the intermediate portion.
From FIG. 5 as well, it can be seen 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.

As described above, in the tapered roller bearing of the present invention, the contact angle α, the ratio D _a / L of the diameter D _a of the large diameter side end of the tapered roller to the length L of the tapered roller, and the roller By limiting the number coefficient k, the dynamic torque can be suppressed to a small value. In order to further reduce this dynamic torque and to prevent seizure even in the case of extremely poor lubrication, in order to improve seizure resistance, the large diameter side end surface of each tapered roller and the large diameter of the inner ring It is effective to limit the surface roughness of one surface of the flange portion formed on the side outer peripheral surface to the surface with which the large-diameter side end surface is in sliding contact to 0.1 μRa or less. Regarding the result of the experiment conducted in order to know the effect of controlling the surface roughness on both sides, FIG.
7 will be described.

First, in an experiment on the influence of the surface roughness on the dynamic torque, a tapered roller bearing having specifications corresponding to B in Table 2 was used, and the oil temperature was 80 at # 90. ℃
900kg of gear oil is supplied at a rate of 600cc / min.
It was performed while applying an axial load of f. 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 is 0.3 μRa, and the solid line b shows the surface roughness of both surfaces. 0.1 μRa
The experimental results of the above are shown. As is clear from FIG. 6, by limiting the surface roughness of both surfaces to 0.1 μRa or less, the dynamic torque in the practical rotation range can be significantly reduced. In the experiments whose results are shown in FIGS. 4 to 5, the roughness of both surfaces was 0.
The measurement was performed by limiting the amount to 1 μRa or less.

Next, FIG. 7 shows the effect of the surface roughness of both surfaces on seizure resistance. An experiment on seizure resistance was carried out with a # 9 tapered roller bearing to be tested.
0.2 cc of gear oil of 0 was attached and rotated at 4000 rpm while applying an axial load of 1000 kgf.
It was performed by measuring the time until the image sticking occurred. In FIG. 7 showing the results of this experiment, the vertical axis represents the time (seconds) until the occurrence of seizure, and the horizontal axis represents the root-mean-square roughness [= {(surface roughness of the large-diameter side end surface of the tapered roller) on both surfaces. ² + (Surface roughness of the surface on one side of the flange where the large diameter side end surface is in sliding contact) ² }
^1/2 ] are respectively represented. Of the plurality of points shown in FIG. 7, three ▲ marks show the experimental results regarding the tapered roller bearings whose surface roughness on both surfaces is regulated to 0.1 μRa or less. The experimental results of tapered roller bearings having a surface roughness smaller than that are shown. As is clear from the description of FIGS. 6 to 7, by restricting the surface roughness of both surfaces to 0.1 μRa or less, the dynamic torque in the practical rotation range is significantly reduced and seizure resistance is improved. I can do things.

As described above, the tapered roller bearing for supporting the pinion shaft of the differential gear of 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, it is possible to secure practically sufficient values for fatigue life and bearing rigidity. The fatigue life and the bearing rigidity can be calculated by knowing the bearing specifications as shown in Table 2 above.

FIG. 8 shows 6 of A to F described in Table 2 above.
The results obtained by calculating the fatigue life of various types of tapered roller bearings are shown. The vertical axis represents the fatigue life of the conventional A and D tapered roller bearings as 1, and the life of the other tapered roller bearings as a ratio to the life of these A and D tapered roller bearings. As is apparent from FIG. 8, the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention is
Fatigue life equivalent to that of conventional products can be secured.

Next, FIG. 9 shows A to F described in Table 2 above.
6 types of tapered roller bearings, combination of A and D, B
With the combination of E and E, and the combination of C and F, the pinion shaft 2 (FIG. 1) is mounted at two positions in the front and rear of the middle part, and the pinion shaft 2 is rotatably supported.
The results obtained by calculating the axial bearing rigidity are shown. The vertical axis shows the axial bearing rigidity (displacement amount) when the pinion shaft 2 is supported by the conventional A and D tapered roller bearings, and the axial bearing rigidity when other tapered roller bearings are combined. It is described as a ratio to the axial bearing rigidity in the case of combining both A and D tapered roller bearings. Since the value on the vertical axis is represented as the ratio of the displacement amount when a predetermined axial load is applied, the lower the 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, the axial bearing rigidity equal to or higher than that of the conventional case can be obtained.

Next, FIG. 10 shows the results of calculating the radial bearing rigidity under the same conditions as in FIG. Figure 9 above
Similarly to the case, since the value on the vertical axis is represented as the ratio of the displacement amount when a predetermined radial load is applied, the lower the value, the higher the bearing rigidity. As is clear 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,
The radial bearing rigidity is lower than in the conventional case. Therefore, when both the B and E tapered roller bearings are combined, it can be applied to the differential gear of a vehicle equipped with a high-power engine as in the case of combining the conventional A and D both tapered roller bearings. When combined with a tapered roller bearing, it can be applied to a differential gear of an automobile equipped with a low-power engine.

By considering the results of the many experiments including the experiments described above, the present inventor has determined that the contact angle is 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 unique 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 the pinion shaft of a differential gear. We came to think that the action and effect could be obtained. Therefore, as shown in G1 to G8 in Table 3 below, regarding the above-mentioned three types (the above-mentioned) limiting elements, eight types of samples in which the above-mentioned upper limit value and lower limit value are respectively combined are taken, and When the performance required for the tapered roller bearing for supporting the pinion shaft was measured, it was found that sufficient performance could be obtained.

[0028]

[Table 3]

From the above description and the description in Table 3 above, the scope of the present invention is that a parallelepiped (for example, a rectangular parallelepiped or a cube) having eight points represented by G1 to G8 in Table 3 as its vertices. You can see that it is represented inside. Also, these G1 to G8
The surface roughness of the large diameter side end surface of each tapered roller and the surface of the flange portion formed on the large diameter side outer peripheral surface of the inner ring with which the large diameter side end surface is in sliding contact with It was also found that if all of them are regulated to 0.1 μRa or less, good results can be obtained as in the case of B, C, E and F.

[0030]

Since the tapered roller bearing for supporting the pinion shaft of the differential gear of the present invention is constructed and operates as described above, the power required in the differential gear portion is ensured while ensuring the performance required for the differential gear. It can reduce loss and contribute to fuel efficiency of automobiles.

[Brief description of drawings]

FIG. 1 is a vertical cross-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 vertical cross-sectional view of the experimental device.

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

FIG. 5 is a diagram showing a relationship between rotation speed and 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 influence 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 a ratio of axial bearing rigidity.

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

[Explanation of symbols]

 1 Casing 2 Pinion shaft 3 Coupling flange 4 Reduction small 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 raceway 11a, 11b Inner ring raceway 12 Holder 13 Drive shaft 14 Outer Holder 15 Housing 16 Oil Filling Hole 17 Static Pressure Pad 18 Arm Piece 19 Load Cell

Claims (4)

[Claims] 1. A pinion shaft having a front end portion freely connectable to a rear end portion of a propeller shaft and a reduction small gear meshing with a reduction gear large gear fixed to the rear end, and two positions before and after an intermediate portion of a pinion shaft are rotated with respect to a casing. In a tapered roller bearing for freely supporting a pinion shaft of a differential gear, the contact angle α is 22 to 28 degrees, and the diameter D _{a of the} tapered roller on the large-diameter side is D _a.
The ratio D _a / L of the length L of the tapered roller and is from 0.51 to 1.0
And the diameter of the pitch circle of the plurality of tapered rollers is 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.
A tapered roller bearing for supporting a pinion shaft of a differential gear, wherein the tapered roller bearing has a size of 6 to 1.32. 2. The contact angle α is 26 to 28 degrees and the ratio D _{a is}
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 between the large-diameter side end surface of each tapered roller and the surface of the flange portion formed on the large-diameter side outer peripheral surface of the inner ring with which the large-diameter side end surface is in sliding contact is 0. The tapered roller bearing for supporting a pinion shaft of a differential gear according to any one of claims 1 to 3, wherein the tapered roller bearing has a diameter of 1 μRa or less.