JPH10331866A - Differential sensitive type hydraulic coupling device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent control characteristics from being changed caused by the exothermic of a clutch. SOLUTION: The difference in rotation ΔN of a differential induction type hydraulic coupling device 12 between a first rotating shaft 17 linked with a main driving wheel 2 and a second rotating shaft 19 linked with a driven wheel 11, allows a hydraulic pump 24 to be driven, allows a clutch 25 to be fastened, allows the rotating shafts themselves to be fastened, and also allows the driving force of the main driving wheel 2 to be appropriately distributed to the driven wheel 11. In the aforesaid device in particular, its control characteristics are so designed as to be stable by dislocating a hydraulic control mechanism 90 and the clutch in position in the axial direction, and letting the heat of the clutch be hardly transmitted to the hydraulic control mechanism. Besides, the hydraulic control mechanism is separated from the clutch by partitioning members 37, 38, 48 and 64, and it is so designed that working oil suitable to both of them but different in nature, shall be used. Furthermore however, working oil is commonly used in the clutch 25 and a separate driving mechanism.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a differentially responsive hydraulic coupling device, and more particularly to a differential responsive hydraulic coupling device utilizing a rotational difference between a main drive wheel (for example, a rear wheel) and a sub drive wheel (for example, a front wheel) of a vehicle.
The present invention relates to a differential-sensitive hydraulic coupling device that switches the driving state of a vehicle from 2WD to 4WD.

[0002]

2. Description of the Related Art When a main driving wheel directly driven by a driving force from an engine slips or idles on a snowy road or a low μ road, the driving force is also transmitted to a secondary driving wheel, that is, an auxiliary driving wheel. A coupling device that switches the driving state of a vehicle from 2WD to 4WD is known. As such a coupling device, the present applicant has previously driven a hydraulic pump using a rotation difference between a main drive wheel and a slave drive wheel, and engaged the drive wheels by fastening a clutch with the generated hydraulic pressure. A differential-sensitive hydraulic coupling device that switches the driving state to 4WD was proposed (Japanese Patent Application No. 8-43634).

In this case, the hydraulic pressure generated by the hydraulic pump is appropriately reduced by a hydraulic control mechanism, and the reduced hydraulic pressure is applied to a piston to control the clutch engagement force. And the rise characteristic of the transmission torque with respect to the rise of the drive wheel rotation difference keeps a constant value after rise in square proportion,
It has a merit that it is easy to handle, heat generation and power loss are small, and the structure is light and simple. Further, since the hydraulic control mechanism is disposed at the axis of the rotating shaft, the influence of centrifugal force due to rotation is eliminated, and stable pressure regulation performance can be secured.

[0004]

By the way, in the previous proposal, since the wet multi-plate clutch is arranged radially outside the hydraulic control mechanism, the operating oil of the hydraulic control mechanism is heated by the heat generated by the clutch. However, there has been a problem that the control characteristic change based on the viscosity change is remarkable.

[0005] Further, since the hydraulic oil for controlling the pressure by the hydraulic control mechanism and the hydraulic oil for the clutch are common, it is not preferable in terms of function. That is, when controlling the pressure, it is preferable to use a smooth operating oil having a low viscosity, that is, an operating oil having a small change in viscosity with respect to a temperature change. On the other hand, it is preferable to use a high-viscosity hydraulic oil that can withstand use at a relatively high temperature for lubricating the clutch. Therefore, it is not appropriate to share these functions to satisfy both functions. In addition, the use of the common pressure is not preferable because the hydraulic oil to be pressure-controlled is heated by the clutch.

[0006]

SUMMARY OF THE INVENTION The present invention is a differential-sensitive hydraulic cup for appropriately distributing a driving force transmitted to a main drive wheel to a sub-drive wheel based on a rotation difference between the main drive wheel and a sub-drive wheel. A ring device, which is driven by a first rotation shaft linked to a main drive wheel, a second rotation shaft linked to a slave drive wheel, and a rotation difference between these rotation shafts, and generates a hydraulic pressure according to the rotation difference. A hydraulic pump, a hydraulic control mechanism for appropriately reducing the hydraulic pressure generated by the hydraulic pump, and fastening of the rotating shafts based on the hydraulic pressure sent from the hydraulic control mechanism, and an axial position with respect to the hydraulic control mechanism And a displaced clutch.

According to this, since the positions of the hydraulic control mechanism and the clutch in the axial direction are shifted, the heat of the clutch is less likely to be transmitted to the hydraulic control mechanism, and the heating of the control hydraulic oil is prevented and stable. Control characteristics can be obtained.

Here, it is preferable that the hydraulic control mechanism and the clutch are separated from each other by a partition member. In this case, different hydraulic oils suitable for each of them can be used. Further, it is preferable that the clutch also serves as another drive mechanism and hydraulic oil.

[0009]

Preferred embodiments of the present invention will be described below in detail with reference to the accompanying drawings.

FIG. 3 shows the entirety of a vehicle to which the differential-sensitive hydraulic coupling device according to the present invention is applied. As shown in the drawing, the vehicle 1 is an FR-based vehicle having a rear wheel 2 as a main drive wheel. It is a 4WD car. That is, the driving force generated by the engine 3 is transmitted to the transfer 6 via the clutch 4 and the transmission 5, and further transmitted to the rear wheel 2 via the propeller shaft 7 for the rear wheel, the differential 8, and the drive shaft 9 for the rear wheel. Is transmitted. The engine 3 and the rear wheel 2 are directly connected, and the driving force is constantly transmitted to the rear wheel 2.

Particularly, in the transfer 6, the driving force to the front wheels 11 is taken out via the chain mechanism 10. However, this driving force will be selectively transmitted to the front wheels 11 by the hydraulic coupling device 12 adjacent to the transfer 6, which will be described later in detail. When the driving force is output from the hydraulic coupling device 12, the driving force is transmitted to the front wheels 11 via the front wheel drive shaft 13, the front wheel differential 14, and the front wheel drive shaft 15. Thus, the front wheels 11 form auxiliary driving wheels or auxiliary driving wheels that are driven auxiliary according to the running state of the vehicle. The front wheel 11 is also a steering wheel, and the chain mechanism 10 can be replaced with a gear or a belt mechanism.

FIG. 1 is a longitudinal sectional view showing details of the hydraulic coupling device 12, and FIG. 2 is an enlarged view of a main part thereof.

As shown in FIG. 1, the hydraulic coupling device 12 has a substantially cylindrical and compact structure as a whole. Specifically, the hydraulic coupling device 12 has a first rotating shaft 17 disposed on the axial center side and a second rotating shaft 19 disposed on the radially outer side thereof. Are rotatably connected to each other through a pair of bearings 18 and 2 which are entirely mounted on a part of the transfer case 16.
It is rotatably supported via a zero. That is, the first rotating shaft 17 and the second rotating shaft 19 form a coaxial double shaft that can rotate relative to each other. The first rotating shaft 17 is connected to the chain 2 of the chain mechanism 10 in the transfer case 16.
The driving force of the engine 3 is obtained from the sprocket 1 and the sprocket 22, and is linked with the rear wheel 2. On the other hand, as shown in FIG.
3 is connected to the front wheel 11 via a universal joint 23.

As will be described later in detail, when the front and rear wheels 11 and 12 are rotating at the same speed in a normal running state of the vehicle, the first rotating shaft 17 and the second rotating shaft 19 are also the same. It rotates in the same direction at the speed and there is no rotation difference. However, for example, when the rear wheel 2 is given an excessive driving force and slips while traveling on a snowy road, the first rotating shaft 17 rotates at a higher speed than the second rotating shaft 19, and a rotation difference occurs. The hydraulic coupling device 12 drives the hydraulic pump 24 using the rotation difference, engages the clutch 25 with the generated hydraulic pressure, distributes or outputs the driving force to the front wheels 11, and changes the driving state to 2WD. (FR) is switched to 4WD. Conversely, if there is no rotation difference, the hydraulic pump 24 is not driven and the clutch 25 is not engaged, so that the driving state is 2WD. As described above, since 2WD is usually used, improvement in fuel efficiency and the like due to low friction can be achieved, and 4WD is automatically set only when necessary, so that operation can be facilitated and safety can be improved.

Next, the configuration of such an apparatus will be described in detail.

The first rotating shaft 17 is mainly composed of a cylindrical shaft 30 and a hollow shaft 31 fixed to the shaft center thereof. The cylindrical shaft 30 is entirely housed in the transfer case 16, and the left half of the hollow shaft 31 is fitted to the shaft center of the cylindrical shaft 30. The key 34 prevents relative rotation, and the fixing nut 33 prevents axial movement. It is fixed. The sprocket 22 is keyed around the outer periphery of the cylindrical shaft 30, and the chain 21 is wound around the sprocket 22. A stopper member 3 is provided at the left end of the cylindrical shaft 30 to prevent the cylinder shaft 30 from coming off.
9 is fixed. A seal member 35 is provided at the right end of the transfer case 16 for sliding contact with the second rotating shaft 19, and a hydraulic piston 74 for pushing a rod 72 is provided at the left end.
Is provided. In particular, when the shift lever is operated to the reverse stage, the hydraulic pressure is applied to the hydraulic piston 74. Based on the hydraulic pressure, the hydraulic piston 74 moves rightward from the illustrated position (see FIG. 10), and To the right.

FIG. 2 shows the second rotating shaft 19 in detail.
Is a large-diameter cylindrical shaft 40 supported by the bearings 26 and 20 from the inner and outer peripheral sides, a cylindrical member 41 fitted into the large-diameter cylindrical shaft 40 from the right side, and a right side joined to the cylindrical member 41 and And a joint member 42 projecting therefrom, which are integrally connected by bolts 27. Cylindrical member 4
A bearing support member 28 is sandwiched between the bearing member 1 and the joint member 42, and this bearing support member 28 is
The right end of the hollow shaft 31 of the first rotating shaft 17 is rotatably supported. The right end opening of the joint member 42 is a flexible partition 43
And the cover 44. Flexible partition 43
Is formed by bending a thin metal plate or an elastic plate such as rubber into a bellows shape. The lid body 44 is provided with the air hole 32, so that the flexible partition wall 43 communicates with the outside.

The cylindrical member 41 extends leftward in the large-diameter cylindrical shaft 40, and has, at its left end, a protruding portion 36 protruding radially inward. On the other hand, a reaction force receiving ring 37 is fitted to the outer peripheral portion of the hollow shaft 31 at the radially inner side of the protruding portion 36 so as to be rotatable but not movable in the axial direction. further,
A seal member 38 that seals a gap between the reaction force receiving ring 37 and the projecting portion 36 is provided on the outer peripheral portion.

The reaction force receiving ring 37 has a right end protruding to the right of the seal member 38 and a cutout 58 provided at the right end thereof, which is provided intermittently in the circumferential direction. Notch 58
Communicates with the hydraulic port 45 of the hollow shaft 31 so that the hollow shaft 31
The inside and outside are communicated with each other. In addition, in the hollow shaft 31,
A plurality of other hydraulic ports 46, 47a, 47b are bored.

The hydraulic pump 24 is disposed coaxially with the large-diameter cylindrical shaft 40 at a position on the right side of the shaft, and has the same configuration as that proposed by the present applicant in Japanese Patent Application No. 6-277790. Radial piston pump. First, the principle will be described.

FIG. 4 is a vertical sectional front view similar to FIG. 2, and FIG. 5 is a vertical sectional side view. Radial piston pump a has a rotation axis b of the center and O _1, on the outer circumference of the rotation shaft b is eccentric portion c are formed integrally. The center O ₂ of the eccentric portion c is offset by S with respect to the center O ₁ of the rotation axis b. A housing d is fitted to the rotation shaft b from the outside so as to be relatively rotatable. As shown in FIG. 5, the inner surface of the housing d and the outer surface of the eccentric portion c are formed in a regular polygonal cross section.
A piston e is pressed to the outside by a pair of rings f and is fixed to each flat surface portion of the inner surface. On the other hand, outside the eccentric portion c, a cylinder ring g shown in FIG. 6 is provided so as to be able to slide and rotate with respect to the eccentric portion c. The cylinder ring g accommodates the eccentric portion c in the center hole h, and accommodates each piston e in the cylinder hole i in the outer peripheral portion so as to be slidable reciprocally. As a result, the cylinder ring g cannot rotate relative to the housing d.

As shown in FIG. 5, if the housing d is fixed and only the rotating shaft b rotates in the direction indicated by the arrow, an eccentric circle in which the cylinder ring g moves only in the radial direction as the eccentric portion c rotates. Do exercise. When this happens, cylinder bore i
Moves reciprocally relative to the piston e, thereby causing a change in the volume of the cylinder chamber j, and the suction and discharge of the hydraulic oil are sequentially performed. In the illustrated state, the hydraulic oil is finally discharged from the port k in a region A where the next discharge is performed. The region B is a region where the next suction is performed, and the hydraulic oil is sucked from the port l. The moving direction of the hydraulic oil is in the direction of the dashed arrow in FIG.

On the other hand, if the rotation direction of the rotation shaft b is opposite to the direction indicated by the arrow, the area A is on the suction side and the area B is on the discharge side. Then, the moving direction of the hydraulic oil in FIG. 4 also becomes as indicated by the solid arrow.

Here, even when the housing d is rotatably supported from the outside, the same operation can be obtained if a relative rotation, that is, a rotation difference occurs between the housing d and the rotation shaft b.
That is, such a radial piston pump a
Is driven by the rotation difference between the housing d and the housing d to generate hydraulic pressure. Conversely, when there is no rotation difference, the rotation shaft b and the housing d only rotate together and are not driven, No hydraulic pressure is generated.

Returning to FIGS. 1 and 2, in the hydraulic pump 24 based on this principle, the hollow shaft 31 corresponds to the rotary shaft b, and the housing d corresponds to the housing d. Is a cylindrical member 41. The eccentric portion c corresponds to the eccentric portion 50 formed integrally with the hollow shaft 31, and the cylinder ring g corresponds to the eccentric portion 50. The piston 52 corresponds to the piston e, and the ring 53 corresponds to the ring f.

The eccentric portion 50 has separate rings 54 and an integral ring portion 55 at the left and right ends.
The ring portion 55 sandwiches the cylinder ring 51 and the pair of rings 53, thereby restricting the movement in the axial direction. As shown in the upper part of the figure, a balancer 56 is provided integrally with the ring 54 and the ring portion 55. A plurality of pistons 52 are provided in the circumferential direction as described above,
Each cylinder hole 57 of the cylinder ring 51 is slidably fitted in a reciprocating manner. The cylinder ring 51 is fitted to the outside of the eccentric part 50 so as to be relatively rotatable.

Next, the clutch 25 has a wet multi-plate clutch structure. That is, the clutch 25
Consists of a plurality of clutch plates 60 and 61 arranged alternately. The inner clutch plate 60 is on the cylindrical shaft 30, the outer clutch plate 61 is on the large-diameter cylindrical portion 40, and the spline 6
Attached via the two and 63, it is movable in the axial direction and is not movable in the circumferential direction.

To the right of these clutch plates 60 and 61,
A hydraulic piston 64 for executing engagement / disengagement (or connection / disconnection) of the clutch 25 is provided. The hydraulic piston 64 has a U-shaped cross section that opens rightward, and an inner surface thereof forms a hydraulic working surface 65. The tip of the reaction force receiving ring 37 is in sliding contact with the inner surface of the outer peripheral wall 66 of the hydraulic piston 64, and the inner peripheral wall 67 of the hydraulic piston 64 is in sliding contact with the hollow shaft 31. As a result, as will be described in detail later, the leakage of the hydraulic pressure acting on the working surface 65 is prevented, and a hydraulic chamber 68 is defined on the right side of the working surface 65.

When hydraulic pressure is applied to the operating surface 65, the hydraulic piston 64 is pushed to the left, and the clutch plates 60,
61 are pressed against each other to fasten (connect) the clutch 25. The reaction force at this time is received by the reaction force receiving ring 37 abutted against the ring 54. Then, when the oil pressure is lost, the pressing force of the clutch plates 60 and 61 is also lost, and the clutch 25 is automatically released (separated).

By the way, such a hydraulic coupling device 1
2 mainly includes a hydraulic piston 64 and a reaction force receiving ring 3.
7 and a seal member 38 oil-tightly partition the inner peripheral side from the substantially middle position in the axial direction to the left and right. The right end of the right side region is closed by a flexible partition 43 and a lid 44. Further, at the shaft center, a seal pin 48 fitted with an O-ring is fitted inside the hollow shaft 31 to partition the inside oil-tightly left and right. Thus, the oil chamber 70 in which the control hydraulic oil is sealed is formed on the right side inside the device, and the clutch chamber 49 in which the clutch lubricating hydraulic oil can enter is formed in the outer peripheral portion where the left clutch 25 is located. It will be. Here, the gear oil in the transfer case 16 is also used as the hydraulic oil for clutch lubrication, and this oil is used for the communication port 59 provided in the large diameter cylindrical shaft 40 and the communication oil provided in the cylindrical shaft 30 and the hollow shaft 31. Port 116, 1
17 penetrates into the clutch chamber 49. The seal pins 48 are arranged on the right side of the rod 72.
The communication ports 116 and 117 are bored in the radial direction to supply oil to the clutch chamber 49 using centrifugal force generated by rotation of the cylindrical shaft 30 and the hollow shaft 31.

As will become clear later, a hydraulic control mechanism 90 for controlling the clutch engagement force is immersed in the oil chamber 70, and the hydraulic piston 64 and the reaction force receiving ring 3 described above are immersed in the oil chamber 70.
The seal member 38 and the seal pin 48 constitute a partition member for separating the hydraulic control mechanism 90 and the clutch 25 from each other.

The control hydraulic oil filled in the oil chamber 70 is
A hydraulic oil having a small change in viscosity with respect to a temperature change and suitable for control is used. The hydraulic oil is sucked by the hydraulic pump 24 and discharged in a high pressure state.
And the clutch 25 is engaged. The hydraulic pressure generated by the hydraulic pump 24 is sent to the clutch 25 through a hydraulic passage 71 along the axial direction inside the hollow shaft 31. Since the right end of the oil chamber 70 is partitioned by a flexible partition 43, and the flexible partition 43 can be freely deformed while breathing with the outside through the air holes 32, the inside of the oil chamber 70 is normally maintained at the atmospheric pressure. Becomes The flexible partition 43 can be replaced with a free piston.

As shown in detail in FIG. 2, a substantially right half of the hollow shaft 31 is enlarged in diameter, and a control sleeve member 75 is disposed in a central hole 78 in the hollow shaft 31 by fitting. Further, a stopper 69 is fixed to a right end of the control sleeve member 75 with a nut 118 after the slide ring 73 is fitted.
Further, between these and the right end of the hollow shaft 31, a port ring 76 and an end ring 77 are arranged in the axial direction and fitted like a sealing member. These are prevented from coming off collectively by the stopper ring 89, and as will become clear later, the movement in the axial direction is completely restricted or limited to a predetermined amount. Note that movement in the rotation direction is allowed. These have hydraulic ports 79, 80, 81, 82, respectively.
83, 84 and 85 are formed. Further, black portions in the drawing indicate seals such as O-rings, whereby each sliding contact portion is oil-tightly sealed.

The control sleeve member 75 is urged to the left by a spring 87, but its movement is restricted by the slide ring 73 abutting against the port ring 76. A stopper member 88 is provided at the left end of the control sleeve member 75, and a center port 86 communicating with the external oil chamber 70 is provided at the right end.

In particular, inside the control sleeve member 75,
A hydraulic passage 71 for sending the hydraulic pressure generated by the hydraulic pump 24 to the clutch 25 is formed. And each hydraulic port 4
.. Form a communication passage that connects the hydraulic pump 24, the hydraulic chamber 68 or the oil chamber 70 to the hydraulic passage 71. Various members described below are inserted into the hydraulic passage 71,
These members constitute a hydraulic control mechanism 90 for appropriately reducing the hydraulic pressure generated by the hydraulic pump 24 and applying the reduced pressure to the clutch 25.

As shown here, the hydraulic control mechanism 90
The clutch 25 and the clutch 25 are displaced from each other in the axial direction, that is, the hydraulic control mechanism 90 is disposed on the right side of the position of the hydraulic piston 64, and the clutch 25 is disposed on the left side of the hydraulic piston 64. This makes it difficult for the heat generated by the clutch 25 to be transmitted to the hydraulic control mechanism 90, so that extremely stable control characteristics can be obtained.

The hydraulic control mechanism 90 is provided with a sleeve body 97 fitted into the substantially left half of the control sleeve member 75.
A first valve body 91 and a second valve body 92 coaxially and doubly fitted in the sleeve body 97, and a flow-sensitive piston 98 fitted on the right side of the sleeve body 97;
Further, it is mainly constituted by a third valve element 93 disposed on the right side thereof. The first and second valve bodies 91 and 92 are urged to the right by first and second springs 94 and 95, respectively.
These springs 94 and 95 are arranged coaxially and in a fitted state. Flow rate sensitive piston 98 and third valve element 9
The third spring 96 is arranged in a compressed state between the third spring 96 and the third spring 96.
As a result, the flow rate sensitive piston 98 is urged to the left and abuts against the second valve body 92, and the third valve body 93 is urged to the right and abuts the right end inner wall of the control sleeve member 75.

The sleeve body 97 has a hydraulic port 99 extending in the axial direction and a hydraulic port 100 drilled in the radial direction. On the right side of the hydraulic port 100, a second valve hole 101 is provided.
The second valve body 92 is fitted in the second valve hole 101. The second valve hole 101 is formed in a tapered shape in which the right side is reduced in diameter, and the fitting portion of the second valve body 92 is also formed in a tapered shape in accordance with this. A substantially left half portion of the sleeve body 97 defines a spring chamber 102 to form the first first part.
And the second springs 94 and 95. The spring chamber 102 is communicated with the outside by a hole provided in the stopper member 88.

The second valve body 92 is slidably fitted to the sleeve body 97 from the left side.
5 for receiving the urging force. The second valve body 92 is entirely formed in a cylindrical shape extending in the axial direction, and the first valve body 91 is slidably fitted therein. A tubular projection 104 having a predetermined inner diameter is formed at the right end of the second valve body 92, and the tubular projection 104 projects to the right of the second valve hole 101 and abuts on the flow-sensitive piston 98. The tubular projection 104 defines a throttle passage 114 at the axial center. Further, in the second valve body 92, a tapered first valve hole 105 whose diameter is also reduced to the right is formed at the same position as the second valve hole 101 in the axial direction. The first valve hole 105 communicates with the throttle passage 114, and the needle portion 106 of the first valve body 91 having a tapered tip is fitted into the first valve hole 105. A hydraulic port 107 is bored on the left side of the first valve hole 105 in the radial direction.

The first valve body 91 is slidably fitted to the second valve body 92 from the left, and has a seat 108 at the left end for receiving the urging force of the first spring 94. The first valve body 91 is formed as a solid shaft extending entirely in the axial direction, and has a needle portion 106 at the right end thereof. The needle portion 106 has a cylindrical portion whose second valve body 9
There is a gap between the inner surface and the inner surface 2 to allow the passage of hydraulic oil.

The first valve body 91 is pushed by the first spring 94 to close the first valve hole 105. In this state, the second
Since the spring 95 also presses the second valve body 92, the valve bodies 91 and 92 close the second valve hole 101. As will be described in detail later, when the first valve body 91 is moved to the left by receiving the hydraulic pressure, the movement is eventually restricted by the stopper member 88, and the stroke is limited to a predetermined amount. Also, in this state, when the second valve body 92 is moved to the left by receiving the hydraulic pressure, the seat portions 103 and 108 come into contact with each other and the movement is restricted, and the stroke is limited.

Next, the flow rate sensitive piston 98 is entirely formed in a cylindrical shape extending in the axial direction, and allows the passage of hydraulic oil at its axial center. Further, a throttle portion 109 projecting radially inward is formed on the inner surface of the throttle portion 109. As will be described in detail later, a thrust is generated by a differential pressure between the upstream side and the downstream side of the throttle portion 109, and the pressing force on the left side is reduced. Is to occur. A hydraulic port 110 is formed at the left end of the flow rate sensitive piston 98 along the radial direction.

The third valve element 93 is also formed in a cylindrical shape extending entirely in the axial direction. Hydraulic ports 111 and 112 communicating with the hydraulic ports 81 and 82 of the control sleeve member 75 are provided on the peripheral wall portion. Can be An orifice hole 113 having a very small diameter is provided on the right end wall. As will be described later in detail, the orifice hole 113 forms an oil discharge port to constantly release the hydraulic pressure in the hydraulic passage 71 to the external oil chamber 70.

Next, the operation of the hydraulic coupling device 12 having the above configuration will be described.

FIG. 11 is a graph showing the torque transmission characteristics of this device. The horizontal axis represents the rotational difference ΔN between the front and rear wheels 11 and 12.
However, the transmission torque T to the front wheels 11 is shown on the vertical axis.
Here, the rotation difference ΔN is used as it is for the first and second rotating shafts 1.
It corresponds to the rotation difference of 7, 19, that is, given by ΔN = (the rotation speed of the rear wheel 2) − (the rotation speed of the front wheel 11), and accordingly, ΔNa = (the rotation speed of the first rotation shaft 17) − (Second rotating shaft 1
9). The transmission torque T is
5 according to the magnitude of the hydraulic pressure applied to the hydraulic piston 26.

As shown in the figure, the torque transmission curve is divided into areas A to D, which will be described in detail later. In the region A, the transmission torque T increases from 0 in proportion to the square of the rotational difference ΔN to T _1. In the region B, the transmission torque T is constant at T ₁ even if the rotation difference ΔN increases, and in the region C, the transmission torque T decreases as the rotation difference ΔN increases.
A region D is a region where the rotation difference ΔN is negative, and here, T = 0
It has become. Note that the rotation difference ΔN that defines these regions
Is 0 between DA, N ₁ between AB, and N ₂ between BC.

Hereinafter, these areas will be described in order.
First, a relatively small rotation difference Δ due to slip of the rear wheel 2 or the like.
In the area A where N occurs, the hydraulic coupling device 12 is in the state shown in FIG. 1 (or FIG. 2), and the movement of the hydraulic oil is performed as shown by the arrow in the figure.

Referring to FIG. 2 as well, the hydraulic pump 24 is supplied with the low-pressure hydraulic oil in the oil chamber 70 through the first port 130 of the eccentric part 50.
Inhaled from. Here, the first port 130 is communicated with the oil chamber 70 via the hydraulic ports 47b, 85, 84. On the other hand, high-pressure hydraulic oil is discharged from the second port 131 located at a position facing the first port 130,
a, 83, 81, and 111 are sent to the hydraulic passage 71 to pressurize the inside of the hydraulic passage 71. Further, the hydraulic pressure propagates to the left in the hydraulic passage 71, and the hydraulic ports 110, 99, 7
9 and 46, the oil is transmitted to the hydraulic chamber 68 in order, and pushes the hydraulic piston 64 to the left, thereby engaging the clutch 25.

On the other hand, when the inside of the hydraulic passage 71 is pressurized, a part of the hydraulic oil in the hydraulic passage 71 is discharged from the orifice hole 113 which is the only conductive path to the oil chamber 70. Due to the throttle effect at the time of passing through the orifice hole 113, the hydraulic pressure on the hydraulic passage 71 side increases in a square proportional to the hydraulic oil flow rate. The hydraulic oil flow rate is
, That is, a value proportional to the rotation difference ΔNa between the first and second rotating shafts 17 and 19, and eventually the hydraulic passage 71
The hydraulic pressure inside increases in proportion to the square of the rotation difference ΔN between the front and rear wheels. Then, as shown in a region A of FIG. 11, the transmission torque T also becomes square proportional to the rotation difference ΔN.

As described above, the third valve body 93 having the orifice hole 113 constitutes an orifice member for determining the characteristic of increasing the hydraulic pressure applied to the clutch 25 as described above.

Next, in the region B where the slip of the rear wheel 2 increases and the rotation difference becomes N ₁ ≤ΔN≤N ₂ , the device 12
7 and the movement path of the hydraulic pressure are as shown in FIG.

As shown in the drawing, when the hydraulic pressure reaches a value (first predetermined value) corresponding to the transmission torque T ₁ , the spring constant of the first spring 94 is smaller than that of the second spring 95. In response to this, first, the first valve body 91 moves to the left, and the first valve hole 105 is opened with an appropriate throttle amount. In this case, in addition to the orifice hole 113,
Hydraulic ports 107, 100, 80, 45 and notch 58
, The oil pressure is discharged to the oil chamber 70.
Then, the first valve hole 10 is added to the passage area of the orifice hole 113.
As a result of the addition of the throttle passage area formed by the first valve body 5 and the first valve body 91, the throttle effect at the time of hydraulic oil flowing out to the oil chamber 70 is reduced. rise rate is reduced small, the transmission torque T is to be held (torque limit) at a constant value T _1. Further, the first valve body 91 increases the amount of movement as the hydraulic pressure increases, and enables the discharge of the hydraulic pressure according to the rotation difference. When the first valve element 91 is the stopper member 88
If it abuts, its movement is regulated.

As described above, the first valve hole 105 forms an oil discharge port which is opened and closed by the movement of the first valve body 91. The first valve hole 105 and the first valve body 91 are connected to each other by the hydraulic pump 24.
Opens when the hydraulic pressure generated at reaches the first predetermined value,
A first relief valve that suppresses an increase in hydraulic pressure applied to the clutch 25 is configured.

Next, in a region C where the rotation difference ΔN satisfies N ₂ ≦ ΔN, the state of the device 12 and the moving path of the hydraulic pressure are as shown in FIG.

As described above, when the first valve hole 105 is opened in the region B, a flow of hydraulic oil from the outlet of the hydraulic port 111 toward the first valve hole 105 is generated in the hydraulic passage 71.
In the course of this flow there is a flow-sensitive piston 98, all of which passes through the throttle 109 of this flow-sensitive piston 98. At the time of this passage, a differential pressure proportional to the flow rate of the hydraulic oil is generated between the upstream side (pump side) and the downstream side (the first valve hole side) of the throttle section 109, and this differential pressure eventually causes the flow rate sensitive piston 98 Thrust to the left.

On the other hand, since the second valve body 92 itself also receives the hydraulic thrust toward the left side, when the sum of the differential pressure thrust and the hydraulic thrust exceeds the set load of the second spring 95, the second The two-valve element 92 starts moving leftward against the second spring 95, and starts opening the second valve hole 101.
The hydraulic pressure at this time is a hydraulic pressure generated when the rotation difference ΔN = N ₃ . When the second valve hole 101 is opened, the hydraulic oil passing through the second valve hole 101 is discharged to the oil chamber 70 through the hydraulic ports 107, 100, 80, 45 and the notch 58, and the hydraulic passage 71
The hydraulic pressure inside the machine starts to drop. Further, the second valve body 92 increases the amount of movement to the left as the generated hydraulic pressure increases, and the second valve hole 1
01 gradually increases the opening amount. Also, as the release amount increases, the discharge flow rate of the hydraulic oil also increases, and accordingly, the differential pressure thrust of the flow rate sensitive piston 98 also increases. After all, due to these cooperative actions, in the region C, the rotation difference ΔN
As the hydraulic pressure increases, the hydraulic pressure acting on the hydraulic piston 64 keeps decreasing, and the engagement force of the clutch 25 and the transmission torque T also continue to decrease.

Here, the passage formed by the second valve hole 101 and the second valve body 92 is located radially outside the passage formed by the first valve hole 105 and the first valve body 91. The former passage area is larger than the latter passage area, and therefore the rotation difference ΔN
The increase rate of the passage area and the decrease rate of the hydraulic pressure with the increase of the hydraulic pressure are larger in the former than in the latter. In the initial stage of movement of the second valve body 92, the first valve body 91 is pushed to the left by hydraulic pressure to open the first valve hole 105. However, if the hydraulic pressure increases to some extent and the second valve body 92 moves by a predetermined amount, The seats 103 and 108 come into contact with each other,
As if the second valve body 92 and the first valve body 91 are integrated,
At the same time, the first valve hole 105 closes. In this state, the movement of the valve bodies 92 and 91 to the left is caused by the two springs 9.
The movement is limited by the total urging force of 4, 95, and the movement is allowed until the first valve body 91 hits the stopper member 88.

As described above, the second valve body 92 and the second valve hole 1
01 indicates that the hydraulic pressure generated by the hydraulic pump 24 is equal to the rotation difference N ₂
The second relief valve is configured to open when the pressure reaches a second predetermined value corresponding to, and to reduce the hydraulic pressure applied to the clutch 25. In the above-described region C, the transmission torque is reduced with respect to an insignificant increase in the rotation difference after completing the function of guaranteeing the necessary maximum transmission torque in the region B, and excessive heat generation in the clutch 25 is reduced. This is the target area to prevent.

Next, the area D will be described. This region is a region where the negative rotation difference ΔN is opposite to the above-described positive rotation difference ΔN.
This is a region where the rotation shaft 19 rotates at a higher speed than the first rotation shaft 17. In practice, this corresponds to a case where the turning radius of the front wheels 11 increases during turning of the vehicle, resulting in high rotation.

As shown in FIG. 9, in this case, the hydraulic pump 24 is driven in the reverse direction, and the lower first port 130 in FIG. 9 is on the discharge side, and the upper second port 131 is on the suction side. Thus, the suction side and the discharge side of the ports 130 and 131 are switched or switched.

The hydraulic oil discharged from the first port 130 is discharged to the oil chamber 70 through the hydraulic ports 47b, 85, 84.

On the other hand, immediately after the rotation difference ΔN becomes negative, the hydraulic oil in the hydraulic passage 71 is sucked into the second port 131 on the suction side through the hydraulic ports 111, 81, 83 and 47a. Due to this suction, a negative pressure is generated in the hydraulic passage 71, and the hydraulic oil in the oil chamber 70 is supplied through the orifice hole 113. However, the suction amount by the hydraulic pump 24 is much larger than the suction amount by this. The inside of the hydraulic passage 71 is in a negative pressure state. At this time, since the negative pressure also acts on the hydraulic piston 64, the engagement force of the clutch 25 is completely released, and the vehicle 1 is completely in the 2WD state. If such a negative pressure state is excessive, a problem such as cavitation occurs, which is not preferable. Therefore, the third valve element 93 is provided for the purpose of preventing the problem.

That is, the third valve body 93 always has the hydraulic passage 7
1, the third valve element 93 is pressed to the right when the oil pressure is in a positive pressure state. For example, as shown in FIG. The portion abuts on the right end taper portion inside the control sleeve member 75 to partition between the oil chamber 70. However, when the hydraulic pressure becomes negative, as shown in FIG.
Since a force acting to draw the third valve body 93 to the left side acts, when this force exceeds the set load of the third spring 96, the third valve body 93 moves to the left side, separating the contact portion thereof, and A communicating portion with the oil chamber 70 is formed between the member 75 and the member 75.

Therefore, if the set load of the third spring 96 is set to be small, before the generated negative pressure becomes excessive,
Hydraulic oil can be immediately sucked into the hydraulic passage 71 through the communication section and the hydraulic ports 86 and 112, thereby solving problems such as cavitation.

As described above, when the second rotary shaft 19 rotates at a higher speed than the first rotary shaft 17, the third valve body 93
A check valve that is opened by the negative pressure in 1 and communicates the hydraulic passage 71 and the oil chamber 70 is configured.

Next, when the vehicle is moving backward, the following is performed.
Referring to FIG. 1, first, when the driver operates the shift lever to the reverse gear, the hydraulic piston 74 is pressurized from the left in conjunction with this, and accordingly, the rod 72 and the seal pin 48 are pushed to the right. Become like

Then, as shown in FIG. 10, the entire control sleeve member 75 slides rightward in the hollow shaft 31, and the slide ring 73, whose movement is restricted by the end ring 77, moves leftward on the control sleeve member 75. Move relative to. In this case, the hydraulic port 82 communicates with the hydraulic ports 85 and 47b, the first port 130 communicates with the hydraulic passage 71, the hydraulic port 81 is shut off, and the hydraulic port 84 is connected with the hydraulic port 8
The second port 131 communicates with the oil chamber 70 by communicating with the third and 47a. Here, the first and second rotating shafts 17,
19, the rotation direction is opposite to that in the case of forward movement, and when the first rotation shaft 17 becomes higher rotation than the second rotation shaft 19, hydraulic oil is discharged from the first port 130, and The suction side and the discharge side of the second ports 130 and 131 are switched.

Accordingly, this hydraulic oil or oil pressure follows the same route as in the case of the forward movement, and generates a fastening force on the clutch 25 to bring the vehicle 1 into the 4WD state.

As described above, the rod 72 and the seal pin 4
Reference numeral 8 denotes a port switching unit that switches a communication state between the first and second ports 130 and 131, the hydraulic passage 71, and the oil chamber 70 in accordance with the forward and backward traveling of the vehicle.

Next, the features of the present apparatus that performs such an operation will be described.

The features of this apparatus are as follows.
The point that the rotating shaft 17, the second rotating shaft 19, and the hydraulic pump 24 are arranged coaxially and the hydraulic control mechanism 90 is arranged on the axis thereof is raised. In this way, the hydraulic control mechanism 90 is not affected by centrifugal force even under high-speed rotation, accurate operations of each member are secured, and stable and accurate torque transmission control can be performed. In addition, there is an advantage that the whole apparatus becomes compact and downsizing is achieved.

Next, as a second feature of the present apparatus, the point that the orifice hole 113 can make the transmission torque T proportional to the square of the rotation difference ΔN is raised.

That is, by doing so, torque transmission when the rotation difference ΔN is minimal is suppressed, and deterioration of fuel efficiency due to heat generation and friction loss can be prevented. FIG. 12 shows the torque transmission characteristics of the conventional viscous coupling device. In this case, excessive torque is transmitted even when a slight differential occurs between the front and rear wheels when traveling on a general uneven road or turning. Would. This apparatus has an advantage that this can be solved.

Next, as a third feature of the present apparatus, the point that the transmission torque T can be maintained at a constant value T ₁ by the first valve body 91 is raised.

In general, the μ characteristics of a multi-plate clutch are as shown in FIG. 13, that is, the friction coefficient μ tends to decrease gradually as the rotation difference ΔN increases. Therefore, as shown in FIG. 14, the present apparatus has a so-called override characteristic in which the supply oil pressure P to the hydraulic piston 64 increases as the rotation difference ΔN increases. This gives the friction coefficient μ
, The clutch engagement force and the transmission torque T can be kept constant. When such a torque limit is performed, it is not necessary to give a great strength to the drive system of the front wheels 11, and the weight can be reduced.

Next, as a fourth feature of the present apparatus, the point that the transmission torque T can be reduced from the fixed value T ₁ by the second valve element 91 is raised.

That is, at the time of the above-described torque limit, in principle, the actual excessive load torque of the drive system is intended to be kept within a certain range by heat conversion of excess excessive torque energy and discarding it. Therefore, in this case, heat is generated in proportion to (transmission torque × rotational difference).

However, as shown in FIG. 15, even after the transmission torque (T) reaches the limit torque (T ₁ ), an operation is performed to further increase the differential rotation (ΔN) by blowing the engine. In such a case, the torque transmission is limited while the differential rotation only increases, which not only results in meaningless fuel consumption, but also generates heat (H) proportional to the differential rotation, resulting in an abnormal rise in temperature. There is a problem.

Therefore, in the present apparatus, after maintaining a constant torque to some extent, the transmission torque is reduced to suppress such heat generation and abnormal rise in temperature, and to improve the durability of the apparatus. I have.

As shown in FIG. 16, when the rotation difference ΔN is N ₁ ≦
In a region B where ΔN ≦ N ₂ , the heat value H increases as the rotation difference ΔN increases. And the region C of the rotation difference ΔN ≧ N ₂
In, to keep the heating value to a value when the .DELTA.N = N ₂ may be caused to coincide with the heating limit torque T _H illustrating the transmission torque T. Therefore, in the present apparatus, the value is set to a value smaller than the above value, so that heat generation and abnormal rise in temperature are completely prevented. As a result, the durability and safety of the device are improved, and the transmission torque is reduced from a constant value, so that the load on the front wheel drive system is reduced, and the weight and cost can be reduced due to the reduced strength.

Next, as a fifth feature of the present apparatus, when the front wheel 11 rotates at a higher speed than the rear wheel 2 by the third valve body 93, the vehicle 1 is moved to 2WD without engaging the clutch 25.
The point that can be maintained is raised.

That is, during turning, a difference in turning locus between the front and rear wheels occurs, and the front wheels rotate at high speed. If the 4WD state is set at this time, a so-called tight corner braking phenomenon occurs, and smooth turning is prevented. In the present device, when the front wheel 11 is rotating at high speed, torque is not transmitted to prevent this.

A sixth feature of the present apparatus is that the vehicle 1 can be brought into the 4WD state when the vehicle is moving backward as in the case of moving forward.

That is, when the rear wheel 2 rotates at a higher speed when the vehicle is moving backward, the hydraulic pump 24 is driven to rotate in the reverse direction, so that the suction side port and the discharge side port are switched. In this state, that is, in the state of FIG. 1, the discharged hydraulic pressure is immediately discharged to the low pressure side, and the clutch 25 cannot be engaged.

Therefore, in the present apparatus, these ports are switched to apply the discharged hydraulic pressure to the clutch 25 so that the front wheels 11 can be driven. As a result, 4WD can be achieved even when the vehicle is moving backwards, and it is possible to easily put the vehicle in a garage on a snowy road.

When the front wheel 11 rotates at a high speed during reverse travel, the third valve element 93 moves in the same manner as described above.
Becomes 2WD, and a smooth turn is secured.

The above features are based on the previously proposed Japanese Patent Application No. Hei 8-43634.
This device also has the following features.

That is, the seventh feature of the present apparatus is that the hydraulic control mechanism 90 and the clutch 25 are disposed so as to be shifted in the axial direction. This makes it difficult for the high heat generated by the frictional sliding of the clutch 25 to be transmitted to the hydraulic control mechanism 90, and does not raise the temperature of the hydraulic oil used for the hydraulic control mechanism 90, thereby avoiding control or malfunction due to a change in viscosity thereof. It becomes possible.

The eighth feature of the present apparatus is that the hydraulic control mechanism 90 and the clutch 25 are separated from each other by a partition member. By doing so, the hydraulic control mechanism 9
A hydraulic oil having a low viscosity for 0 and suitable for control with a small change in characteristics with respect to a temperature change is used. The hydraulic oils can be used individually, and both functions can be sufficiently achieved. In the hydraulic control mechanism 90, stable control characteristics can always be obtained, and in the clutch 25, abnormal vibration such as stick-slip of the sliding surface due to too low viscosity, and abnormal increase in the internal pressure due to the vaporization phenomenon, etc. It can be prevented beforehand.

In particular, since the clutch 25 also serves as another drive mechanism (transfer 6) and hydraulic oil (lubricating oil), the cost can be reduced and the cooling performance can be improved by increasing the amount of oil. Can be achieved. In addition, there is no need to consider sealing due to the use of clutch-specific hydraulic oil.
From such a concept, the present device can be adjacent to another drive mechanism, for example, the transmission 5, the front wheel differential 14, and the like, and can also use the gear oil thereof.

Next, the ninth feature of this device is that
The point that the valve body 91 and the second valve body 92 are arranged in a fitted state with each other is raised. In previous proposals (Japanese Patent Application No. 8-43634, etc.), these are spool valve type valves that are arranged along the axis, which increases the shaft length and the entire device. There are drawbacks. This device prevents this,
There is an advantage that the shaft length can be shortened and the whole device can be made compact. Since the first and second springs 94 and 95 are also provided in a fitted state, the shaft length can be further reduced. Further, the former spool valve type has a disadvantage that the dimensional relationship between the inclined surface of each valve element and the oil discharge port is sensitive, and extremely high processing accuracy is required. By arranging the device as described above, it is possible to form a double poppet valve type valve body.
1 and the tapered opening / closing portions of the respective valve elements 91 and 92 can be easily and accurately processed, and there is an advantage that the manufacturing cost can be reduced.

Further, as a tenth feature of the present apparatus, a point that the spring constant of the first spring 94 is smaller than that of the second spring 95 can be pointed out. Thus, a small capacity relief valve can be configured by the first valve body 91 and a large capacity relief valve can be configured by the second valve body 92. In particular, variation in sensitivity due to variation in the spring constant of the first spring 94 can be suppressed, and a stable torque limit can be achieved. Can be realized.

In particular, in the region A, in the first valve body 91, the pressure is received only at the tip of the needle portion 106 facing the first valve hole 105, and the area of this portion is small. The spring constant of one spring 94 can be reduced. Since the opening / closing portion is formed in a tapered shape, the rate of change in the opening area with respect to the increase in the hydraulic pressure can be set gently, whereby an appropriate override amount matching the μ characteristics of the clutch 25 can be obtained.

By the way, after the first valve body 91 moves to the specified amount, the oil pressure rises in proportion to the square of the rotation difference due to the throttle effect, while the μ characteristic of the clutch 25 saturates as the rotation difference increases. In this state, the override amount is excessively large, which is not preferable.

Therefore, at this time, the second valve element 92 is opened and the hydraulic oil is discharged at a large flow rate, whereby the hydraulic pressure is reduced at a stretch and the above-mentioned torque reduction is performed. As a result, it is possible to prevent the temperature from rising abnormally and to reduce the weight by reducing the strength.

Next, as an eleventh feature of the present apparatus, the point that a flow rate sensitive piston 98 is provided is raised. That is, the flow rate sensitive piston 98 applies a pressing force to the second valve body 92 in accordance with the flow rate of the hydraulic oil passing through the inside. For this reason, in the region C where the large rotation difference ΔN occurs, the second valve body 92 can be opened larger than in the case where only the second valve body 92 is used, whereby a reliable torque reduction can be achieved, and complete protection of the device can be achieved. Etc. become possible.

As for the positional relationship between the first valve element 91 and the second valve element 92, as described above, the first valve element 91 is disposed on the axis side and the second valve element 92 is disposed radially outward. Is preferred. This is because the former can reduce the pressure receiving area and can open the valve with a small pressure, and the latter can increase the pressure receiving area and the passage area, and can open the valve with a large pressure and increase the amount of oil discharge. Therefore, it is preferable that the second valve body 92 is formed in a tubular shape that can accommodate the first valve body 91. Further, they are preferably arranged at the axial center of the apparatus to eliminate the influence of centrifugal force.

On the other hand, various modifications are conceivable for these, for example, the arrangement relationship between the first valve body 91 and the second valve body 92 can be reversed, or a sleeve or the like is interposed between them. It is also possible. Similarly, the arrangement relationship between the first spring 94 and the second spring 95 can be reversed.

Another feature of the present apparatus is that the engagement reaction force of the clutch 25 is given by the axial force of the first rotating shaft 17. That is, like the present apparatus, the first rotating shaft 17
And the second rotating shaft 19 have a double pipe structure that can rotate relatively,
When these are to be fastened by the clutch 25, the rotating shafts tend to separate from each other due to the fastening reaction force, and this causes an excessive thrust force to be applied to the bearing portion, thus causing a problem that the durability of the bearing portion is impaired. Therefore, in this device, the reaction force is received by the axial force of the first rotating shaft 17, thereby preventing the deterioration and damage of the bearing portion. Specifically, referring to FIG. 2, a rightward reaction force when hydraulic pressure is applied to hydraulic chamber 68 is received by step portion 115 of first rotating shaft 17 through reaction force receiving rings 37 and 54. . The leftward reaction force is received by the step 120 of the cylindrical shaft 30 via the clutch positioning ring 119. As a result, it is possible to prevent an excessive thrust force from being applied to the bearings 26 and 29,
The bearings 26 and 29 can be protected.

Further, other features of the present apparatus include the first
The rotary shaft 17 has a divided structure of a cylindrical shaft 30 and a hollow shaft 31, and a subunit is configured by assembling the components individually with each other, and finally these shafts are assembled together to assemble the whole. This facilitates assembly and improves productivity, and also ensures that performance and quality control is ensured. Specifically, the clutch positioning ring 119, the bearing 26, the large-diameter cylindrical shaft 40, and the clutch plates 60 and 61 are mounted on the cylindrical shaft 30 to complete one subunit, and the hollow shaft 31 is mounted with another component. Complete the other subunit. Finally, these units are fitted to each other, and the cylindrical shaft 30 and the hollow shaft 31 are fixed to the fixing nut 3.
If they are connected to each other by the key 3 and the key 34, the entire device can be assembled. In particular, the former sub-unit is a power transmission system, so that its dimensional accuracy and the like are not so strict. However, since the latter sub-unit is a hydraulic control system, its dimensional quality control and performance check must be strictly performed. Therefore, by unitizing each function in this way,
In particular, the latter performance and quality control can be sufficiently performed, and reliable operation can be performed.

Although the preferred embodiment of the present invention has been described above, the present invention is not limited to this embodiment, and can adopt various other embodiments.

[0102]

The present invention exhibits the following excellent effects.

(1) It becomes difficult for the heat generated by the clutch to be transmitted to the hydraulic control mechanism, whereby a change in the viscosity of the control hydraulic oil is prevented and stable control characteristics can be obtained.

(2) Different working oils suitable for the clutch and the hydraulic control mechanism can be used.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional front view showing an entire hydraulic coupling device according to the present invention, and particularly shows a movement path of hydraulic oil in a region A.

FIG. 2 is an enlarged view of a main part of FIG.

FIG. 3 is a configuration diagram of a vehicle to which the hydraulic coupling device according to the present invention is applied.

FIG. 4 is a vertical sectional front view for explaining the principle of the hydraulic pump.

FIG. 5 is a vertical sectional side view of FIG. 4;

FIG. 6 is a perspective view showing a cylinder ring.

7 is a vertical sectional front view of the hydraulic coupling device, particularly showing a moving path of hydraulic oil in a region B. FIG.

8 is a vertical sectional front view of the hydraulic coupling device, particularly showing a movement path of hydraulic oil in a region C. FIG.

FIG. 9 is a vertical sectional front view of the hydraulic coupling device, particularly showing a moving path of hydraulic oil in a region D;

FIG. 10 is a vertical sectional front view of the hydraulic coupling device, particularly showing a moving path of hydraulic oil when the vehicle is moving backward.

FIG. 11 is a graph showing a torque transmission characteristic of the hydraulic coupling device according to the present invention.

FIG. 12 is a graph showing torque transmission characteristics of a conventional viscous coupling device.

FIG. 13 is a graph showing μ characteristics of the multi-plate clutch.

FIG. 14 is a graph showing changes in hydraulic pressure and transmission torque to a clutch.

FIG. 15 is a graph showing a relationship between a transmission torque and a heat value with respect to a rotation difference.

FIG. 16 is a graph showing a relationship between a transmission torque and a heat value with respect to a rotation difference.

[Explanation of symbols]

 2 Rear wheel 11 Front wheel 12 Hydraulic coupling device 17 First rotating shaft 19 Second rotating shaft 24 Hydraulic pump 25 Clutch 37 Reaction force receiving ring 38 Seal member 48 Seal pin 64 Hydraulic piston 90 Hydraulic control mechanism ΔN Rotational difference

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[Procedure amendment]

[Submission date] June 19, 1997

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Claims (3)

[Claims] 1. A differential-sensitive hydraulic coupling device for appropriately distributing a driving force transmitted to a main drive wheel to a sub-drive wheel based on a rotation difference between the main drive wheel and a sub-drive wheel, comprising: A first rotating shaft interlocked with the wheels, a second rotating shaft interlocked with the driven wheels, a hydraulic pump driven by a rotation difference between these rotation shafts, and generating a hydraulic pressure according to the rotation difference; A hydraulic control mechanism for appropriately reducing the hydraulic pressure generated by
A differential-sensitive hydraulic coupling, comprising: a clutch for fastening the rotating shafts based on a hydraulic pressure sent from the hydraulic control mechanism, and a clutch whose axial position is shifted with respect to the hydraulic control mechanism. apparatus. 2. The differential-sensitive hydraulic coupling device according to claim 1, wherein the hydraulic control mechanism and the clutch are separated from each other by a partition member. 3. The differential-sensitive hydraulic coupling device according to claim 2, wherein the clutch also serves as another drive mechanism and hydraulic oil.