JPH09236142A - Differential sensing type hydraulic coupling device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To set a region for slow rising of transmission torque and a region for constant transmission torque by providing a plurality of relief valves for gradually reducing hydraulic pressure given to a clutch. SOLUTION: A hydraulic port 81 and the first tapered surface 117 of a first valve element 91 are opened when a hydraulic pressure produced by a hydraulic pump reaches a first prescribed value equivalent to a rotation difference so as to form a first relief valve for reducing the increase rate of a hydraulic pressure given to a clutch. The hydraulic port 81 and the second tapered surface 118 of the first valve element 91 are opened when a hydraulic pressure produced by the hydraulic pump reaches a second prescribed value equivalent to a rotation difference so as to form a second relief valve for further reducing the increase rate of the hydraulic pressure given to the clutch.

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, a differential-sensing type that uses a rotation difference between a main driving wheel and an auxiliary driving wheel is generally used, and a viscous cup that uses a highly viscous fluid is widely used at present. There is a ring device.

On the other hand, a hydraulic coupling device for driving a hydraulic pump by utilizing the rotation difference of the drive wheels, engaging a clutch with the generated hydraulic pressure to connect the drive wheels, and switching the drive state to 4WD is fair. It is known from 5-57131. In this case, when the main drive wheels do not slip and there is no rotation difference between the drive wheels, the hydraulic pump is not driven and the clutch is not engaged, so the vehicle is in the complete 2WD state. Therefore, the automatic 4WD is realized in which only the main drive wheels are slipped when the slip or the like occurs, and the 4WD is set only when necessary.

In particular, in the case of the former viscous coupling device, although the rise of the transmission torque is gentle and easy to handle in the normal region, the rise of the transmission torque is rapid in the region where the rotational difference is extremely small, so that the road surface can be generally used. Even if a difference in rotation occurs due to unevenness of the road surface during traveling, excessive torque is frequently transmitted, and heat generation loss causes deterioration of fuel efficiency. In addition, the tight corner braking phenomenon becomes noticeable even when turning. Further, if the rotation difference becomes large due to the depression of the accelerator pedal, a high torque is transmitted to the auxiliary drive wheels, so that the drive system is also required to have a corresponding strength.

On the other hand, in the latter hydraulic coupling device, as shown in Japanese Patent Publication No. 57131/1993, a part of the hydraulic pressure discharged from the hydraulic pump is discharged from the orifice to square the rotational difference. Proportional torque transmission characteristics can be obtained, the rise of the transmission torque can be suppressed, particularly in the region where the rotation difference is extremely small, and the above-mentioned problems can be solved. Further, if a relief valve is provided in the hydraulic circuit system, a torque limiter that suppresses an increase in the transmission torque will be realized, which will prevent high torque transmission to the auxiliary drive wheels and reduce the weight of the drive system. There are merits.

[0006]

However, when the transmission torque rises in proportion to the square of the rotation difference as described above, the rise of the torque is gentle in the minute rotation difference region, and the drawbacks of the viscous coupling device can be solved. The transmission torque suddenly rises due to the rotation difference, and the gradual torque increase characteristic, which was advantageous in the viscous coupling device, is lost. For example, the advantage that is advantageous when starting up a hill is lost. .

Further, in the torque limit as described above, in principle, excess energy of excessive torque is converted into heat and discarded, so that heat generation proportional to (transmission torque × rotational difference) is generated, resulting in rotation difference. There is a problem that the temperature rises abnormally due to the increase of.

[0008]

SUMMARY OF THE INVENTION The present invention is a differential-sensing hydraulic cup that appropriately distributes a driving force transmitted to a main driving wheel to a sub driving wheel based on a rotation difference between the main driving wheel and the sub driving wheel. A ring device that is driven by a first rotating shaft that interlocks with a main driving wheel, a second rotating shaft that interlocks with a subordinate driving wheel, and a rotation difference between these rotating shafts, and generates hydraulic pressure according to the rotation difference. A hydraulic pump, a clutch that engages the rotating shafts with each other based on the hydraulic pressure from the hydraulic pump, and an orifice that discharges a part of the hydraulic pressure generated by the hydraulic pump and determines the rising characteristic of the hydraulic pressure applied to the clutch. And a first relief valve that opens when the hydraulic pressure generated by the hydraulic pump reaches a first predetermined value to reduce the increase rate of the hydraulic pressure applied to the clutch, and the hydraulic pressure generated by the hydraulic pump. Is greater than the first predetermined value It opens when it reaches a second predetermined value,
A second relief valve that further reduces the increase rate of the hydraulic pressure applied to the clutch, and opens when the hydraulic pressure generated by the hydraulic pump reaches a third predetermined value that is higher than the second predetermined value, and the clutch is released. And a third relief valve for reducing the hydraulic pressure applied to the.

In the above structure, since the first relief valve and the second relief valve reduce the increase rate of the hydraulic pressure in two stages, the torque transmission characteristic can be set in two stages as well, whereby the transmission torque rises. A gentle region and a constant region can be set. Therefore, the gradual torque increase characteristic, which is advantageous in the viscous coupling device, can be obtained, and at the same time, the torque limit can be realized.

Further, since the hydraulic pressure of the third relief valve is reduced, the transmission torque and the amount of heat generation can be reduced, and the temperature rise can be suppressed.

[0011]

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

FIG. 3 shows an entire vehicle to which the differential-sensing hydraulic coupling device according to the present invention is applied. As shown in the figure, the vehicle 1 is an FR-based vehicle having rear wheels 2 as main driving wheels. It is a 4WD vehicle. 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 through 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 vertical cross-sectional view showing the 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 generally cylindrical and compact structure as a whole. Specifically, in the hydraulic coupling device 12, a hollow first rotating shaft 17 is rotatably supported by a part of a transfer case 16 via a bearing 18, and a second rotating shaft is provided inside the first rotating shaft 17. A shaft 19 is inserted and supported by a bearing 20 so as to be relatively rotatable. That is, the first rotating shaft 17
The second rotary shaft 19 and the second rotary shaft 19 form a coaxial double shaft rotatable relative to each other. The first rotating shaft 17 receives the driving force of the engine 3 from the chain 21 and the sprocket 22 of the chain mechanism 10 in the transfer case 16, and the rear wheel 2
It is linked with. On the other hand, the second rotary shaft 19
Is connected to the front-wheel drive shaft 13 via a universal joint 23 so as to interlock with the front wheels 11.

Here, as will be described in detail later, when the front and rear wheels 11, 2 are rotating at the same speed in a normal vehicle traveling state, the first rotating shaft 17 and the second rotating shaft 19 are also the same. Rotate in the same direction at speed and no difference in rotation occurs. 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 structure of such an apparatus will be described in detail.

The first rotary shaft 17 has a cylindrical housing 30.
And a hollow shaft 31 integrally covered on the outside thereof. The housing 30 has a large-diameter portion 32 forming a substantially right half thereof exposed to the outside, and a reduced diameter portion 33 forming a substantially left half thereof is inserted in the transfer case 16. The diameter-reduced portion 33 has a shape in which the diameter is reduced stepwise toward the left end, and the hollow shaft 31 is provided outside the diameter-reduced portion 33.
Are covered and are integrally connected by the key 34.
The aforementioned sprocket 22 is keyed to the outer peripheral portion of the hollow shaft 31, and the chain 21 is wound around the sprocket 22. The transfer case 16 is a pair of bearings 1.
The hollow shaft 31 is supported at both ends in the longitudinal direction via 8 and the entire first rotating shaft 17 is rotatably supported. At the right end of the transfer case 16, a seal member 35 that slidably contacts the hollow shaft 31.
Is provided, and the rod insertion hole 36 is provided at the axial center position on the left end surface. In particular, the rod insertion hole 36 allows a rod (not shown) that advances and retracts in conjunction with the shift lever to be inserted.

The hollow shaft 31 is formed in a substantially stepped shape following the shape of the reduced diameter portion 33, except for the left end portion keyed to the reduced diameter portion 33 and the right end portion fixed only by fitting. Forms a gap 37 with the reduced diameter portion 33. In particular, a peripheral groove 38 having a U-shaped cross section is provided at the right end of the reduced diameter portion 33, and an O ring (not shown) is inserted into the peripheral groove 38.
Although similar peripheral grooves are drawn in other members in the drawing, description thereof will be omitted below, assuming that all O-rings are inserted in these peripheral grooves. A stopper member 39 for preventing the hollow shaft 31 from coming off is fixed to the left end portion of the hollow shaft 31.

As shown in detail in FIG. 2, the second rotary shaft 19 is
Hollow central shaft 40 arranged at the axial center position of the large diameter portion 32
And a support member 41 and a joint member 42 that are integrally fixed to the left and right ends of the central shaft 40, respectively. The support member 41 and the joint member 42 extend outward in the radial direction and are supported by the bearing 20 from the outside. And joint member 4
2 projects outside from the right side open end of the large diameter portion 32 and is connected to the universal joint 23 with a bolt or the like not shown. At the right end of the large-diameter portion 32, a seal member 43 that slidably contacts the joint member 42.
Is provided. The support member 41 is provided with notches 44 intermittently along the circumferential direction, and the notches 44 communicate with the hydraulic ports 45 of the central shaft 40 so that the inside and outside of the central shaft 40 communicate with each other. In addition, the central shaft 40 is provided with the hydraulic port 4
5, a plurality of hydraulic ports 46, 47a along the radial direction,
47b, 48, 49 are drilled.

The hydraulic pump 24 is arranged coaxially with the large-diameter portion 32 at the right side position in the large-diameter portion 32, which has the same structure as that proposed by the applicant in Japanese Patent Application No. 6-277790. This is a 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 rotation axis b rotates in the direction of the arrow, the cylinder ring g moves only in the radial direction as the eccentric portion c rotates. Do exercise. When this happens, the cylinder hole 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 rotating shaft b is opposite to the direction of the arrow, the area A is the suction side and the area B is 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, a similar 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.

Referring back to FIG. 2, in the hydraulic pump 24 based on this principle, the central shaft 40 corresponds to the rotating shaft b, and the large diameter corresponds to the housing d. The part 32. Further, the eccentric portion c corresponds to the eccentric ring member 50 press-fitted and fixed to the central shaft 40,
The cylinder ring 5 corresponds to the above cylinder ring g.
It is one. The piston 52 corresponds to the piston e.
The ring 53 corresponds to the ring f.

The eccentric ring member 50 has protrusions 54 and 55 at the left and right ends thereof, and the protrusions 54 and 55 sandwich the cylinder ring 51 and the pair of rings 53 so as to sandwich them in the axial direction. It regulates movement. As shown in the lower part of the figure, a balancer 56 is integrally attached to the protrusions 54 and 55. Plural pistons 52 are provided in the circumferential direction similarly to the above, and are fitted in the respective cylinder holes 57 of the cylinder ring 51 so as to be reciprocally slidable. Further, the piston 52 is formed in a hollow shape to reduce the weight. The cylinder ring 51 is relatively rotatably fitted to the outside of the eccentric ring member 50.

Next, the clutch 25 has a wet multi-plate clutch configuration. That is, the clutch 25
Is composed of a plurality of clutch plates 60 and 61 arranged alternately. The inner clutch plate 60 is on the central shaft 40, the outer clutch plate 61 is on the large diameter portion 32, and the splines 62 and 6 are respectively.
It is attached via 3, and is movable in the axial direction but immovable in the circumferential direction.

To the left of these clutch plates 60 and 61,
A hydraulic piston 64 for executing engagement / disengagement (or discontinuity) of the clutch 25 is arranged. The hydraulic piston 64 is formed in a U-shape whose cross section opens to the left, and its inner surface forms a hydraulic pressure acting surface 65. In addition, the inner surface of the outer peripheral wall 66 of the hydraulic piston 64 has a support member 4
1 is in sliding contact with the inner peripheral wall 6 of the hydraulic piston 64.
Reference numeral 7 is in sliding contact with the central shaft 40. Although this will be described in detail later, this prevents the hydraulic pressure acting on the working surface 65 from leaking, so that a hydraulic chamber 68 is defined on the left side of the working surface 65.

When a hydraulic pressure is applied to the working surface 65, the hydraulic piston 64 is pushed to the right and the clutch plates 60,
61 are pressed against each other to fasten (connect) the clutch 25. When the hydraulic pressure is lost, the pressing force of the clutch plates 60 and 61 is also lost and the clutch 25 is automatically released (disengaged).
Is done.

By the way, such a hydraulic coupling device 1
In No. 2, low-pressure or atmospheric-pressure hydraulic oil is sealed in the housing 30. Then, this hydraulic oil is sucked by the hydraulic pump 24 and discharged in a high pressure state, causing the hydraulic piston 64 to be pushed and engaging the clutch 25. Therefore, the housing 30 defines the oil chamber 70 that stores the hydraulic oil supplied to the hydraulic pump 24, and the hydraulic pressure generated in the hydraulic pump 24 is a hydraulic passage formed in the central shaft 40 and extending in the axial direction. It is sent to the clutch 25 through 71.

Referring to FIG. 1, the left end of the housing 30 or the left end of the reduced diameter portion 33 is in an open state,
The open portion is oil-tightly closed by the piston member 72. The piston member 72 is slidable along the inner surface of the reduced diameter portion 33, and normally, as shown in the drawing, the stopper ring 7
It is located at the left end that abuts on 3. However, only when the driver operates the shift lever to the reverse gear, the piston member 72 is pushed by the rod (not shown) and moves to the right side (see FIG. 11). Reference numeral 74 is a ball with which the rod abuts.

A control rod 75 is arranged to the right of the piston member 72, and the control rod 75 is always in contact with the piston member 72 via a ball 76. As shown in FIG. 2, the control rod 75 has a hollow portion 77 forming a substantially right half portion thereof which is axially slidably inserted into a central hole 78 inside the central shaft 40. In particular, the hydraulic passage 71 is defined inside the hollow portion 77, and the right end portion of the central hole 78 is connected to the outside through a hydraulic port 49 to form an oil chamber 70.

In particular, a plurality of hydraulic ports 79, 80, 81, 8 spaced apart in the axial direction are provided on the peripheral wall portion of the hollow portion 77.
2, 83, 84 and 85 are formed so as to penetrate in the radial direction. The hydraulic ports 79 ... and the hydraulic ports 46 of the central shaft 40 form a communication passage that connects the hydraulic pump 24, the hydraulic chamber 68 or the oil chamber 70 to the hydraulic passage 71.
Then, various members described below are inserted into the hydraulic passage 71, and the hydraulic pump 24 is inserted by these members.
A hydraulic pressure control mechanism 90 is configured to appropriately reduce the hydraulic pressure generated in 1 and apply it to the clutch 25.

The hydraulic control mechanism 90 mainly comprises a first valve body 91, a second valve body 92 and a third valve body 93 which are arranged in order from the left side. These first, second and third valve bodies 9
1, 92, 93 are first, second and third springs 94, 9
It is urged to the right at 5,96. The movement of the first valve body 91 to the right is restricted by the stopper ring 97, and the second valve body 9
The movement of 2 to the right is restricted by the control sleeve 98.

The control sleeve 98 is inserted inside the right end of the hollow portion 77, is biased to the left side by the spring 99, and is pressed and fixed to the hollow portion 77. The spring 99 simultaneously biases the control rod 75 and the piston member 72 to the left via the control sleeve 98. Third valve body 93
Is axially movably inserted in the control sleeve 98 and normally closes the communication port 100 in the middle portion of the control sleeve 98 as shown. And communication port 1
A peripheral wall portion 101 of 00 is located between the hydraulic ports 84 and 85, and hydraulic ports 102 and 103 that communicate the inside and outside of the control sleeve 98 are formed on the left and right sides of the peripheral wall portion 101. In particular, the third valve body 93 and the peripheral wall portion 101 partition the hydraulic passage 71 and the oil chamber 70.

The second valve body 92 is formed in a tubular shape that moves in the axial direction like a spool valve, and an inner surface thereof has a throttle portion 104 for applying hydraulic pressure for movement.
A hydraulic port 105 that communicates the inside and outside of the second valve body 92 is formed on the left side of the throttle portion 104.

The first valve body 91 has an enlarged diameter portion 111 at the right end of the hollow shaft portion 110. Particularly, the hollow shaft portion 110 is inserted into the small diameter port 112 of the control rod 75 so as to be slidable in the axial direction. ing. The small diameter port 112 communicates with the oil chamber 70 at its left end. The left end wall of the hollow shaft portion 110 is provided with an orifice hole 113 having an extremely small diameter, and the intermediate portion is provided with a hydraulic port 114 for communicating the inside and the outside. Expanded part 1
11 is slidable in the axial direction along the inner wall of the hollow portion 77, and normally closes the hydraulic port 81 as shown.

In FIG. 7, the shape of the outer peripheral surface of the expanded diameter portion 111 is shown enlarged. In particular, with reference to FIG. (A), the enlarged diameter portion 111 has an annular surface 115 that comes into full surface contact with the inner wall of the hollow portion 77, and a tapered surface 1 formed on the right side of the annular surface 115.
16 (slope). Especially tapered surface 116
Has a gentle slope on the left first tapered surface 117 and a steep slope on the right second tapered surface 118. That is, the tapered surface 116 is formed in multiple steps (two steps here).

Returning to FIG. 1, the reduced diameter portion 3 of the housing 30.
3 accommodates a free piston 120 which is slidable along its inner peripheral surface. The free piston 120 seals the working oil together with the housing 30 and the piston member 72 to define the oil chamber 70. The free piston 120 is formed into a stepped tubular shape, and the oil chamber 70 is defined by the inner surface of the free piston 120 so that the free piston 120 can move freely as the working oil expands and contracts.

That is, the atmospheric pressure chamber 121 is defined by the outer surface portion of the free piston 120 and the inner surface portion of the reduced diameter portion 33, and the atmospheric pressure chamber 121 is connected to the reduced diameter portion 33 and the communication port 12 of the hollow shaft 31.
It communicates with the inside of the transfer case 16 via 2, 123. As described above, the transfer device using the chain mechanism 10 is housed in the transfer case 16, and thus the transfer case 16 becomes the housing chamber 124 communicating with the atmospheric pressure chamber 121. In particular, normally, the transfer case 16 is provided with a breather (not shown) that communicates the inside with the outside, whereby the accommodation chamber 124 and the atmospheric pressure chamber 121 both become atmospheric pressure, and the oil chamber 70 is also kept at atmospheric pressure. To be done.

The atmospheric pressure chamber 121 is provided in the housing 30.
It becomes a chamber on the opposite side of the oil chamber 70 with the free piston 120 as a boundary, and the breather forms a communicating portion with the outside.

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

FIG. 17 is a graph showing the torque transmission characteristic of such a device, in which the horizontal axis indicates the rotation difference ΔN between the front and rear wheels 11, 2.
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 drawing, the torque transmission curve is divided into regions A to E, which will be described in detail later. In the region A, the transmission torque T increases from 0 to T ₁ in proportion to the square of the rotation difference ΔN. In the region B as well, the transmission torque T increases from T ₁ to T ₂ as the rotation difference ΔN increases, but the rate of increase is smaller than in the region A. In the region C, the transmission torque T is constant at T ₂ even if the rotation difference ΔN increases, and in the region D, the transmission torque T decreases as the rotation difference ΔN increases. Region E is a region where the rotation difference ΔN is negative, and T = 0 here. The rotation difference ΔN dividing these regions is 0 between EA, N ₁ between AB, N ₂ between BC, and N ₃ between CD.

Hereinafter, these areas will be described in order. First, a minimum rotation difference ΔN due to slip of the rear wheel 2 or the like.
In the area A in which the oil pressure occurs, the hydraulic coupling device 12 is in the state of FIG. 1, and the movement of the hydraulic oil is performed as shown by the arrow in the figure. The solid line shows the high voltage side and the broken line shows the low voltage side.

Referring also to FIG. 2, the low pressure hydraulic oil in the oil chamber 70 is drawn into the hydraulic pump 24 from the first port 130 of the eccentric ring member 50. That is, the first port 130 is connected to the oil chamber 70 via the hydraulic ports 48 and 85. 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, 84 to the hydraulic passage 71,
It will move to the left in 1. Further, this hydraulic oil or hydraulic pressure moves to the inside of the first valve body 91 because each hydraulic port is closed, and from the orifice hole 113, which is the only communication path between the hydraulic passage 71 and the oil chamber 70, to the hydraulic port. It is discharged to the oil chamber 70 via 79.

At this time, due to the throttling effect of the hydraulic oil passing through the orifice hole 113, the hydraulic pressure on the hydraulic passage 71 side rises in proportion to the flow rate of the hydraulic oil.

The working oil flow rate is the working amount of the hydraulic pump 24,
That is, the rotation difference ΔNa between the first and second rotating shafts 17 and 19
Therefore, the hydraulic pressure in the hydraulic passage 71 eventually increases in proportion to the square of the rotation difference ΔN between the front and rear wheels.

Since the hydraulic oil in the hydraulic passage 71 is electrically connected to the hydraulic chamber 68 via the hydraulic ports 114, 80, 45 and the notch 44, the hydraulic pressure in the hydraulic passage 71 remains in the hydraulic chamber 68. Is also propagated, and the hydraulic piston 64 is pushed to the right by a force proportional to the hydraulic pressure to engage the clutch 25, so that the transmission torque T becomes squarely proportional to the rotation difference ΔN, as in the region A of FIG. .

As described above, the first valve body 91 having the orifice hole 113 forms an orifice member that determines the rising characteristic of the hydraulic pressure applied to the clutch 25 as described above.

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

That is, when the hydraulic pressure reaches a value corresponding to the transmission torque T ₁ , the first valve body 91 receives the force of the hydraulic pressure to move the first valve body 91 to the first position.
It moves to the left against the spring 94 and moves to the hydraulic port 81.
It comes to open while squeezing. In this case, in addition to the orifice hole 113 as a hydraulic outlet, the hydraulic port 8
Since the route leading to the oil chamber 70 via 1, 46 is also formed, the throttle area of the hydraulic discharge passage is equal to the throttle passage area formed by the hydraulic port 81 and the first valve body 91 in the passage area of the orifice hole 113. As a result, the throttling effect in the hydraulic oil outflow to the oil chamber 70 is alleviated, the increase rate of the hydraulic pressure reaching the clutch 25 is reduced, and therefore the increase rate of the transmission torque T is also reduced. It will become smaller.

FIG. 7B shows the details at this time. As shown in the figure, when the first valve body 91 moves to the left, the hydraulic port 81 comes into contact with the first tapered surface 117 and opens. At this time, since the distance between the hydraulic port 81 and the first taper surface 117 is short, the inlet of the hydraulic port 81 serves as a throttle flow path, and the hydraulic oil flowing into the hydraulic port 81 is appropriately limited to the discharge amount. Will be appropriately limited. Then, if the hydraulic pressure increases as the rotation difference increases, the first valve body 91 moves further to the left, and the narrowed flow path is expanded to increase the hydraulic oil discharge amount. Thus,
The hydraulic pressure can be discharged according to the rotation difference.

As described above, the hydraulic port 81 forms an oil drain port which is opened and closed by the movement of the first valve body 91, and the hydraulic port 81 and the first tapered surface 117 of the first valve body 91 are formed. Is a first relief valve that opens when the hydraulic pressure generated by the hydraulic pump 24 reaches a first predetermined value corresponding to the rotation difference N ₁ and reduces the increase rate of the hydraulic pressure applied to the clutch 25. To do.

Next, when the rotation difference ΔN further increases and the hydraulic pressure increases, the first valve body 91 moves further leftward as shown in FIG. 7 (c), which causes the hydraulic port 81 to move to the second tapered surface 1.
It will open up to 18. At this time, a region C in which the rotation difference ΔN is N ₂ ≦ ΔN ≦ N ₃ is realized, and as will be described later in detail, the pushing force of the hydraulic piston 64 is further reduced, and the pushing force is equal to the rotation difference ΔN. As it increases, the engagement force of the clutch 25 and the transmission torque T become constant. The movement route of hydraulic pressure is the same as in FIG.

Here, in the first valve body 91, since the first tapered surface 117 has a small gradient and the second tapered surface 118 has a large gradient, the discharge amount of hydraulic oil relative to the increase in hydraulic pressure is In the case of the first taper surface 117, it can be made small, and in the case of the second taper surface 118, it can be made large, so that the two-step characteristics can be obtained as described above, and the torque T increases at one of the first taper surfaces 117. The decrease of the rate is controlled by the other second taper surface 118 to increase the torque T (torque limit).
It can be performed. It is also possible to provide the tapered surface 116 with multiple stages to obtain more multi-stage characteristics.

As described above, the hydraulic pressure generated by the hydraulic pump 24 becomes equal to the second predetermined value corresponding to the rotation difference N ₂ between the hydraulic port 81 and the portion of the first valve body 91, especially the second tapered surface 118. When it reaches, the valve is opened to form a second relief valve that further reduces the increase rate of the hydraulic pressure applied to the clutch 25.

Next, in the region D where the rotation difference ΔN is N ₃ ≦ ΔN, the state of the device 12 and the hydraulic pressure movement path are as shown in FIG.

The flow of the hydraulic oil discharged from the hydraulic pump 24 in the regions A, B and C up to this point is sent to the hydraulic passage 71 via the hydraulic ports 47a and 84 and is directed toward the first valve body 91. Although it is flowing, the second valve body 92 is installed in the path of this flow, and all the hydraulic oil passes through the throttle portion 104 of the second valve body 92.

Therefore, the hydraulic pressure in the hydraulic passage 71 is still activated by the action of the throttle portion 104 on the upstream side (pump side) and the downstream side (first valve body side) of the throttle portion 104 of the second valve body 92. Since a differential pressure proportional to the square of the oil flow rate is generated, this differential pressure acts on the second valve body 92 and a thrust force that pushes the second valve body 92 to the left is generated.

When this differential pressure thrust reaches a value exceeding the set load of the second spring 95 that presses the second valve body 92 to the right, the second valve body 92 resists the second spring 95 and moves to the left. Start moving to.

The left side of the second valve body 92, which is determined by the balance between the differential pressure thrust acting on the second valve body 92 due to the hydraulic oil flow rate when the rotation difference ΔN becomes N _3, and the deflection load of the second spring 95. The hydraulic port 82 is provided at a position that opens in the outer peripheral groove provided in the second valve body 92 corresponding to the one-way moving position, and the outer peripheral groove and the downstream side of the hydraulic passage 71 are communicated with the second valve body 92. A port 105 is provided. As a result, when the rotation difference ΔN reaches N ₃ , the area C up to that point
In addition to the orifice hole 113 and the hydraulic port 81, which were the escape routes of the hydraulic pressure in FIG.
The escape route will be added, and the hydraulic pressure will start to drop.

With the decrease in the hydraulic pressure, the first valve body 91 returns to the right by the force of the first spring and the hydraulic port 81 starts to close again, but the opening of the hydraulic port 82 has the opening of the hydraulic port 81. because there is no stop structure by such tapered surface of the rotational difference ΔN because the increased amount of open area of the hydraulic port 82 due to the increase in the rotational difference ΔN is well above the amount of decrease in the opening area of the hydraulic port 81 N ₃ In the area D that exceeds
As the rotation difference ΔN increases, the hydraulic pressure acting on the hydraulic piston 64 continues to decrease, and the engaging force of the clutch 25 and the transmission torque T continue to decrease.

As described above, the region D reduces the transmission torque with respect to the meaningless increase in the rotation difference after the required maximum transmission torque guarantee function in the region C is completed, and the clutch 25
This is an area where the function of the hydraulic pump is essentially unnecessary because it is an area that is intended to prevent excessive heat generation.

Therefore, after the hydraulic port 82 starts to open, the rotation difference further increases and the second valve body moves further to the left, so that the hydraulic passage 71 is formed in the oil chamber 7 on the upstream side of the second valve body.
It is more preferable to provide a hydraulic port 83 that opens to zero. That is, since the hydraulic port 82 is on the downstream side of the throttle portion 104 of the second valve body 92, the hydraulic port 82 is fully opened,
If the rotation difference ΔN continues to increase even after the hydraulic pressure acting on the hydraulic piston 64 has completely dropped, the hydraulic pump 24
The load hydraulic pressure of # starts to rise again in proportion to the square of the rotation difference ΔN due to the influence of the throttle 104. However, if the hydraulic port 83 is opened on the upstream side of the second valve body here, the pump discharge not passing through the throttle 104. Since it serves as an escape path for the oil, not only the hydraulic pressure acting on the hydraulic piston 64 can be lowered, but also the load hydraulic pressure of the hydraulic pump 24 can be maintained at an extremely low value, and the durability of the hydraulic pump 24 is improved.

As described above, the second valve body 92 and the hydraulic ports 82 and 83 open when the hydraulic pressure generated in the hydraulic pump 24 reaches the third predetermined value corresponding to the rotation difference N ₃ , and the clutch is opened. A third relief valve for reducing the hydraulic pressure applied to 25 is formed.

Next, the area E 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.

In this case, as shown in FIG. 10, the hydraulic pump 24 is driven in the reverse direction, and the first pump on the lower side in the drawing is shown.
The port 130 serves as the discharge side and the upper second port 131 serves as the suction side, and the suction side and the discharge side of the ports 130, 131 are switched or switched.

The hydraulic oil discharged from the first port 130 is discharged to the oil chamber 70 from the hydraulic ports 48 and 85.

On the other hand, since the working oil is sucked into the second port 131 on the suction side from the oil chamber 70 through the orifice hole 113, the throttle effect of the orifice hole 113 causes the hydraulic pressure in the hydraulic passage 71 to be increased. Is negative pressure. Therefore, a negative pressure is also applied to the hydraulic piston 64, the engagement force of the clutch 25 is completely released, and the vehicle 1 is in the complete 2WD state. However, in such a hydraulic mechanism, if the pressure of the hydraulic oil becomes an excessively negative pressure state, problems such as cavitation will occur.
A valve body 93 is provided.

That is, the hydraulic passage 7 is always provided in the third valve body 93.
When the hydraulic pressure in 1 acts, and this hydraulic pressure is in a positive pressure state, the third valve body 93 is pressed to the right side, so the oil chamber 70
While the communication port 100 with and is completely closed, when the hydraulic pressure becomes a negative pressure, the third valve body 93 is attracted to the left side. When the force exceeds the set load, the third valve body 93
Moves to the left and the communication port 100 opens.

Therefore, if the set load of the third spring 96 is set to be small, even if the rotational difference in which the hydraulic pump 24 is driven in the reverse direction becomes large, the negative pressure generated by the orifice hole 113 becomes large before the pressure becomes large. Since the hydraulic oil is reasonably sucked through the communication port 100 that opens with an extremely small negative pressure value determined by the set load of the spring 96,
Problems such as cavitation do not occur.

That is, the third valve body 93 is the second side which is on the suction side when the second rotary shaft 19 rotates at a higher speed than the first rotary shaft 17.
A check valve that connects the hydraulic passage 71 and the oil chamber 70 is formed by opening the negative pressure generated by sucking the hydraulic oil in the hydraulic passage 71 from the port 131.

Next, when the vehicle is moving backward, the following occurs.
As shown in FIG. 11, when the driver operates the shift lever to the reverse gear, the rod (not shown) pushes the piston member 72 and the control rod 75 to the right in conjunction with this.

Then, in particular, the hydraulic ports 48 and 84 communicate with each other, the first port 130 communicates with the hydraulic passage 71, the hydraulic ports 47b and 85 communicate with each other, and the second port 131 communicates with the oil chamber 7.
It comes to communicate with 0. Here, the rotation directions of the first and second rotating shafts 17 and 19 are opposite to those in the case of advancing, and when the first rotating shaft 17 becomes higher than the second rotating shaft 19, the first rotation Hydraulic oil is discharged from the port 130,
The suction side and the discharge side of the first and second ports 130 and 131 are switched.

Therefore, this hydraulic oil or hydraulic pressure follows the same route as in the case of forward movement, and the engaging force is generated in the clutch 25 to bring the vehicle 1 into the 4WD state.

As described above, the control rod 75 includes the first and second ports 130 and 131, the hydraulic passage 71 and the oil chamber 7.
A port switching means for switching the communication state with 0 according to the forward and backward movement of the vehicle is formed.

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

The features of this apparatus are as follows:
The point where the rotary shaft 17, the second rotary shaft 19, and the hydraulic pump 24 are coaxially arranged and the hydraulic passage 71 is arranged on the axial center is lifted.

That is, in the hydraulic passage 71, Japanese Patent Publication No.
The hydraulic control mechanism 90 is usually provided as shown in the above-mentioned No., and the hydraulic control mechanism 90 controls the hydraulic pressure applied to the clutch 25 to determine the torque transmission characteristic.

However, in Japanese Examined Patent Publication No. 57131/1993, the hydraulic passage is provided at a position distant from the axial center in the radial direction, which causes each member constituting the hydraulic control mechanism to be affected by centrifugal force.
There is a risk that the operation may be hindered and the hydraulic pressure control or the torque transmission control may not be accurately executed.

In this device, since the hydraulic passage 71 is arranged on the axial center and the hydraulic control mechanism 90 is provided, the centrifugal force does not influence the rotating member even under high speed rotation, and the accurate operation of each member is ensured. As a result, stable and accurate torque transmission control can be executed.

The hydraulic pump 24 has the first and second ports 130 and 131 for suction and discharge on the axial center side, and the hydraulic ports are arranged on the axial center side of the hydraulic pump 24. It is possible to arrange the hydraulic passage 71 on the axis.

Then, the first rotating shaft 17 and the second rotating shaft 19
Since the hydraulic pump 24 and the hydraulic pump 24 are arranged coaxially, there is an advantage that the entire apparatus becomes compact and miniaturization is achieved.

Next, as a second feature of this device, the transmission torque T can be made proportional to the square of the rotation difference ΔN by the first valve body 91, and the rate of increase can be reduced when T ₁ is reached. Is fried.

That is, when the rotation difference ΔN is 0 ≦ ΔN ≦ N ₁ and is very small, the transmission torque T is proportional to the square of the rotation difference ΔN. Torque transmission is suppressed, and deterioration of fuel efficiency due to heat generation and friction loss can be prevented. Further, during cornering when ΔN <0, the transmission torque can be maintained at almost 0, so that a tight corner braking phenomenon can be prevented.

FIG. 18 shows a torque transmission characteristic in a conventional viscous coupling device. As shown in the figure, when the rotational difference ΔN is small, the torque curve rises rapidly and the reverse occurs even when ΔN <0. Symmetrical torque transmission occurs. This causes excessive torque to be transmitted even when there is a slight rotation difference between the front and rear wheels when traveling on a general uneven road, or when there is a differential between the front and rear wheels when traveling around the corner, causing the aforementioned problems. Occurs.

Since the torque transmission characteristic of the area A of the present device is a characteristic proportional to the square of the rotation difference ΔN, in the area B, which is an apparent drive wheel slip area, sufficient transmission torque is ensured while at the minute rotation difference area. Transmission torque can be kept extremely small.

Therefore, the restraint torque is extremely small against a small rotation difference generated when the road surface is traveling on a high μ road surface, or when the road surface is uneven or when the vehicle is turning, which is a problem of a general 4WD vehicle. It is possible to eliminate the forced slip phenomenon of the tire, and to realize an ideal automatic 4WD function that automatically transmits a sufficient assist torque to the auxiliary drive wheels when the main drive wheels slip on a low μ road surface.

Further, the assist torque transmission characteristic to the auxiliary drive wheels when the main drive wheels slip on a low μ road surface is also B
As indicated by the region, the assist torque of the auxiliary drive wheels gradually increases as the rotation difference (slip amount of the main drive wheels) increases. The effect of facilitating control can be obtained by adjusting the accelerator pedal operation of the driver, and a 4WD characteristic that is easy to handle can be obtained.

Next, as a third feature of the present device, the point that the transmission torque T can be maintained at a constant value T ₂ by the first valve body 91 is pointed out.

That is, also in Japanese Examined Patent Publication No. 5-57131, similarly, the hydraulic pressure is released by the relief valve to make the clutch engaging force or the transmission torque constant.

However, generally, the μ characteristic of the multi-disc clutch is as shown in FIG. 19, that is, the friction coefficient μ tends to gradually decrease as the rotation difference ΔN increases. Therefore,
As shown in FIG. 20, even if the oil pressure P applied to the clutch by the relief valve is kept constant, the transmission torque T gradually decreases as the rotation difference ΔN increases, and the transmission torque T decreases.
There is a problem that can not be constant.

As a countermeasure against this, it is conceivable to give the relief valve a so-called override characteristic in which the hydraulic pressure on the high pressure side (inlet side) increases as the relief flow rate increases, and this offsets the reduction of the friction coefficient μ. Thus, the clutch engaging force can be made constant.

However, the relief valve shown in Japanese Examined Patent Publication No. 57131/1993 is of the poppet type, and if the relief valve fully open opening area is small, there is throttle loss, so the hydraulic pressure on the high pressure side squares as the relief flow rate increases. It will increase in proportion, an excessive override amount will be transmitted, and torque exceeding the limit torque will be transmitted. Therefore, in order to avoid this, it is sufficient to increase the total opening area of the relief valve, but since the sensitivity of increasing the relief opening area with respect to the valve opening lift is high, the sensitivity must be increased in order to obtain a proper override amount. The spring constant of the relief spring must be set extremely high in order to lower the.

However, this causes a large variation in the set load of the relief spring with respect to the dimensional tolerance of the component parts, and a large variation in the valve opening pressure of the relief valve, which deteriorates the torque limit accuracy itself. Occurs.

Therefore, in the present device, the first valve body 91 is a spool valve type relief valve that moves in the direction orthogonal to the hydraulic port 81, and the first valve body 91 is moved in the axial direction to move from the radial hydraulic port 81. It is designed to drain hydraulic pressure. By doing so, even if the fully open opening area of the hydraulic port 81 is increased and the spring constant of the first spring 94, which is a relief spring, is set low, the sensitivity of the opening area of the hydraulic port 81 to increase in hydraulic pressure can be lowered. As a result, it is possible to prevent variations in the valve opening pressure due to dimensional tolerances, and it is possible to realize an extremely accurate torque limit. The hydraulic port 81 may be formed obliquely with respect to the radial direction so as to form an angle other than 90 ° with the moving direction of the first valve body 91.

Further, in the present apparatus, the hydraulic port 81 faces the second tapered surface 118 when the valve is opened. Then, when the hydraulic pressure rises, the second tapered surface 118 moves to the left, expands the inlet region of the hydraulic port 81, and appropriately increases the discharge amount (relief flow rate) of hydraulic oil. As a result, the hydraulic oil discharge amount is gradually increased as the hydraulic pressure increases, and as shown in FIG. 20, the hydraulic pressure P is gradually increased with respect to the rotation difference ΔN, and a suitable override characteristic is obtained to make the transmission torque T constant. It becomes possible to hold.

In particular, if the transmission torque T is kept constant and further increase is suppressed as described above, it is not necessary to give the drive system of the front wheels 11 a great strength, and the weight can be reduced.

FIG. 21 shows the relationship between the moving stroke L of the first valve body 91 and the effective opening area A of the hydraulic port 81. FIG. 22 shows the relationship between the discharge flow rate Q of the hydraulic pump 24 and the working pressure P of the clutch 25 on the hydraulic piston 64.

It can be seen from FIG. 21 that the relief area (II) formed by the second taper surface 118 has a larger rate of increase in the effective opening area A than the relief area (I) formed by the first taper surface 117. Further, from FIG. 22, the effective opening area is A ₁
It can be seen that the working pressure P does not increase as the discharge flow rate Q increases with increasing A ₅ . By considering these characteristics together, the above-mentioned torque transmission characteristics can be obtained.

Next, as the fourth feature of this device, it is possible to reduce the transmission torque T from the constant value T ₂ by the second valve body 91.

That is, in the case of the torque limit as described above, in principle, it is intended to suppress the actual load torque of the drive system within a certain range by thermally converting and discarding excess energy of excessive torque. is there. So in this case,
Heat is generated in proportion to (transmission torque x rotation difference).

However, as shown in FIG. 23, even after the transmission torque (T) reaches the limit torque (T ₂ ), the engine is further blown to increase the differential rotation (ΔN). If the torque transmission is limited, only the differential rotation increases, meaningless fuel consumption, and heat generation (H) proportional to the differential rotation occurs and the temperature rises abnormally. There is a problem that ends up.

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. There is.

As shown in FIG. 24, the rotation difference ΔN is N ₂ ≦
In the region C where ΔN ≦ N ₃ , the heat generation amount H increases as the rotation difference ΔN increases. And the area D where the rotation difference ΔN ≧ N ₃
Then, in order to maintain the amount of heat generation at the value when ΔN = N ₃ , the transmission torque T may be matched with the heat generation limit torque T _H shown in the figure. Therefore, in order to prevent the heat generation amount from becoming an abnormal value, the transmission torque T may be set to the heat generation limit torque T _H or less, and in this device, a value particularly smaller than that is used to completely prevent heat generation and abnormal temperature rise. I am trying. 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 characteristic of this device, when the front wheel 11 is rotated at a higher speed than the rear wheel 2 by the third valve body 93, the vehicle 1 is driven by 2WD without engaging the clutch 25.
The point that can be maintained is raised.

That is, in Japanese Examined Patent Publication No. 5-57131, the clutch is engaged regardless of which of the front and rear wheels is rotating at high speed.
The vehicle is set to the 4WD state.

However, with this, when the front and rear wheels turn at a high speed due to a difference in the turning loci of the front and rear wheels during turning, dragging occurs on the rear wheels and a tight corner braking phenomenon occurs, resulting in a smooth turning. This will interfere with the operation, resulting in poor maneuverability and fuel efficiency.

Therefore, when the front wheels 11 are not rotated at a high speed as in this device, the driving force is not applied to the front wheels 11.
The free rotation of the front wheels 11 is permitted, the tight corner braking phenomenon is prevented, smooth turning is ensured, and the deterioration of maneuverability and fuel consumption can be prevented.

Next, as a sixth feature of the present device, the fact that the control rod 75 allows the vehicle 1 to be in the 4WD state even when the vehicle is moving backward is the same as when moving forward.

That is, when the rear wheel 2 is rotated at a higher speed when the vehicle is moving backward, the hydraulic pump 24 is driven in the reverse direction, so that its suction side port and its 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 device, the switching of these ports is performed by the control rod 75, and the suction / discharge of the ports is further replaced as shown in FIG. 11, whereby the discharged hydraulic pressure can be given to the clutch 25 to drive the front wheels 11. I am trying. 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.

Further, also in this case, when the front wheels 11 rotate at a high speed, the third valve body 93 opens the communication port 100 and immediately discharges the hydraulic pressure as described above, so that the vehicle 1 becomes 2WD. A smooth turn is possible.

Next, as the seventh feature of this apparatus, the atmospheric pressure chamber 121 partitioned by the free piston 120 is placed in the accommodation chamber 1.
The points communicating with 24 are fried.

That is, the free piston 120 is installed in the oil chamber 70.
Is maintained at atmospheric pressure to allow thermal expansion or contraction of the hydraulic oil and prevent deterioration of the oil seal due to increase in internal pressure.

However, Japanese Patent Laid-Open Nos. 1-250662 and 3-300
Similarly to Japanese Patent No. 266726, if the atmospheric pressure chamber 121 is directly exposed to the atmosphere and the outer surface portion of the free piston 120 and the inner surface portion of the reduced diameter portion 33 (inner surface of the atmospheric pressure chamber 121) are directly exposed to the outside air, they are rusted. Occurrence of dust, adhesion of dust, and the like occurs, which hinders sliding of the free piston 120 and promotes deterioration of the oil seal and the like.

Therefore, in this apparatus, the outside air is not directly introduced into the atmospheric pressure chamber 121, but the atmospheric pressure lubricating oil mist is introduced into the accommodation chamber 124, and the inner surface of the atmospheric pressure chamber 121 is appropriately moistened to generate rust. In addition, dust adhesion is prevented, and functional deterioration of the free piston 120 is surely prevented.

As described above, the accommodating chamber 124 is opened to the atmosphere via a breather (not shown), and the accommodating chamber can be, for example, the inside of a transmission case accommodating a transmission according to the same concept as described above. .

Now, a modified example will be described.

FIG. 12 shows the tapered surface 1 of the first valve body 91.
The modification about 16 is shown. In this case, the tapered surface 116 is not multi-stepped, but is simply a tapered surface having a constant gradient. Even in this case, by appropriately setting the gradient, it is possible to achieve either the decrease in the torque increase rate corresponding to the region B in FIG. 17 or the constant torque holding corresponding to the region C. Further, instead of the tapered surface, a simple flat inclined surface may be used.

In FIG. 13, the first valve body 91 is formed into a cylindrical shape having a constant diameter, and instead, the inner surface of the hollow portion 77 of the control rod 75 is separated from the first valve body 91 and extends radially outward. An example in which the diameter is expanded in a tapered shape is shown. In this, the first
The left end portion of the valve body 91 and the expanded diameter portion 140 of the control rod 75.
(Expanded portion) and hydraulic port 14 as an oil drain port
1, 81 are respectively provided so as to penetrate in the radial direction. Then, normally, the hydraulic port 141 is closed as shown in FIG. However, when the first valve body 91 moves to the left due to the increase in the hydraulic pressure as shown in FIG.
When 41 is opened, the hydraulic pressure is discharged from the hydraulic ports 141 and 81. Particularly at this time, the hydraulic port 141 faces the tapered surface 142 in a spaced manner, and the outlet thereof is the tapered surface 142.
Since it is appropriately throttled at the (slope), the discharge of hydraulic pressure is appropriately limited, and the same effect as described above is brought about. In this example as well, it is possible to form the tapered surface 142 in multiple stages.

FIG. 14 shows an example in which a groove portion 143 is formed along the moving direction in the hydraulic control portion of the first valve body 91 of FIG. That is, in the above-mentioned example, the outer surface is a conical tapered surface, but if the taper gradient is small, the flow toward the hydraulic port 81 immediately after the valve opening becomes a very small gap flow, which is not desirable for the orifice characteristics. There are cases.

Therefore, as shown in FIG. 15, it is possible to improve the orifice characteristics by causing the working oil to sneak into the hydraulic port 81 also from the groove portion 143.
In this example, four grooves 143 are provided and the cross section is a cross shape. Also, the pin 144 shown in FIG.
Controls the relative rotation of the first valve body 91 with respect to the control rod 75.

FIG. 16 shows the groove portion 1 in the first valve body 91 of FIG.
An example in which 45 is formed is shown. Here, a pair of groove portions 145 are provided on both sides in the circumferential direction of the hydraulic port 81. Similar to the example of FIG. 14, the relative rotation of the first valve body 91 with respect to the control rod 75 is regulated by an appropriate means.

Although the preferred embodiments of the present invention have been described above, the present invention is not limited to such embodiments and can take various other forms.

[0128]

The present invention exhibits the following excellent effects.

(1) A gradual torque increase, a constant torque limit, and a torque decrease can be realized in combination, ensuring easy handling, reliable torque transmission at a high speed difference, and suppression of temperature rise. Become.

[Brief description of 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.

FIG. 7 is an enlarged vertical sectional front view of a main part of the first valve body.

FIG. 8 is a vertical cross-sectional front view of the hydraulic coupling device, showing a moving path of the hydraulic oil particularly in regions B and C.

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;

10 is a vertical cross-sectional front view of the hydraulic coupling device, showing a moving path of the hydraulic oil, particularly in a region E. FIG.

FIG. 11 is a vertical cross-sectional front view of the hydraulic coupling device, showing a moving path of the hydraulic oil, particularly when the vehicle is moving backward.

FIG. 12 is an enlarged vertical front view showing a modified example.

FIG. 13 is an enlarged vertical sectional front view showing a modified example.

FIG. 14 is a perspective view showing a modified example.

FIG. 15 is a vertical sectional side view showing a modified example.

FIG. 16 is a perspective view showing a modified example.

FIG. 17 is a graph showing torque transmission characteristics of the hydraulic coupling device according to the present invention.

FIG. 18 is a graph showing torque transmission characteristics of a conventional viscous coupling device.

FIG. 19 is a graph showing μ characteristics of a multi-plate clutch.

FIG. 20 is a graph showing changes in hydraulic pressure and transmission torque to the clutch.

FIG. 21 is a graph showing the relationship between the stroke of the first valve body and the effective opening area of the hydraulic port.

FIG. 22 is a graph showing the relationship between the discharge flow rate of the hydraulic pump and the hydraulic pressure acting on the clutch.

FIG. 23 is a graph showing a relationship between a transmission torque and a heat generation amount with respect to a rotation difference.

FIG. 24 is a graph showing a relationship between a transmission torque and a heat generation amount with respect to a rotation difference.

[Explanation of symbols]

2 rear wheel (main drive wheel) 11 front wheel (slave drive wheel) 12 hydraulic coupling device 17 first rotary shaft 19 second rotary shaft 24 hydraulic pump 25 clutches 81, 82, 83 hydraulic ports (first, second, second) 3 relief valve) 91 1st valve body (1st, 2nd relief valve) 92 2nd valve body (3rd relief valve) 113 Orifice hole (orifice) ΔN Rotational difference

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Tadamichi Kadomura, 4 Isuzu Motor Co., Ltd., 1 Kashiwara, Tomakomai City, Hokkaido Inside the Hokkaido factory (72) Inventor, Hiromi Kato 4 Isuzu Motor Co., Ltd., 1 Kashiwara City, Tomakomai, Hokkaido Company Hokkaido factory

Claims (1)

[Claims] 1. A differential-sensing hydraulic coupling device for appropriately distributing a driving force transmitted to a main driving wheel to a sub driving wheel based on a difference in rotation between the main driving wheel and the sub driving wheel. A first rotary shaft that interlocks with the wheels, a second rotary shaft that interlocks with the driven wheels, a hydraulic pump that is driven by the rotational difference between these rotary shafts, and that generates hydraulic pressure in accordance with the rotational difference; Generated by the hydraulic pump, a clutch that engages the rotating shafts with each other based on the hydraulic pressure of the, and an orifice that discharges a part of the hydraulic pressure generated by the hydraulic pump and determines the rising characteristic of the hydraulic pressure applied to the clutch. Hydraulic pressure is the first
A first relief valve that opens when reaching a predetermined value to reduce the increase rate of the hydraulic pressure applied to the clutch, and a second predetermined value at which the hydraulic pressure generated by the hydraulic pump is greater than the first predetermined value. When it reaches, the second relief valve that opens to reduce the increase rate of the hydraulic pressure applied to the clutch and the hydraulic pressure generated by the hydraulic pump have reached a third predetermined value that is greater than the second predetermined value. A differentially sensitive hydraulic coupling device comprising a third relief valve which is opened at times to reduce the hydraulic pressure applied to the clutch.