JPH0717287A - Torque distributer for vehicle - Google Patents

Abstract

PURPOSE:To save the energy to drive a torque distributing hydraulic motor in a torque distributor for a vehicle. CONSTITUTION:To distribute the torque of an engine E to a left wheel WL and a right wheel WR at a specific ratio, a hydraulic motor 24 to drive a planetary gear device P annexed to a main differential device D is connected to a swash plate type variable displacement hydraulic pump 20. The rotating speed of the hydraulic pump 20 is proportional to the speed of a vehicle and the swash plate connected to a steering wheel has a tilting angle controlled to be proportional to a steering angle. As a result, the discharge amount of the hydraulic pump 20 is restricted to a value that is necessary and sufficient to drive the hydraulic motor 24 at a rotating speed to perform proper torque distribution.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a torque distribution device for a vehicle having a torque distribution drive source for distributing and controlling engine torque to at least two output shafts at a predetermined ratio.

[0002]

2. Description of the Related Art A differential device provided in a power transmission system of an automobile is configured to absorb a rotational speed difference generated between left and right wheels during turning of the automobile and distribute an engine torque to the left and right wheels at an appropriate ratio. To be done. However, since a general differential device operates due to the difference in load applied to the left and right wheels, when one wheel rides on a road surface having a small friction coefficient and spins, the amount of torque transmitted to the other wheel is reduced. There is a problem of reduction or interruption of torque transmission.

In order to avoid such inconvenience, a torque distribution device for a vehicle which positively controls a differential device based on a steering angle and a vehicle speed and distributes a torque suitable for a driving state at that time to left and right wheels. It is proposed in Japanese Patent Laid-Open No. 1-282075. According to this torque distribution device, it is possible to control the magnitude of the torque distributed to the left and right wheels by the hydraulic motor that operates with the pressure oil discharged by the hydraulic pump.

[0004]

However, the above-described conventional torque distribution device for a vehicle is used when torque is not distributed (that is, when the hydraulic motor is not operating) while the vehicle is traveling. However, since the hydraulic pump discharges the pressure oil, energy for driving the hydraulic pump is wasted.

The present invention has been made in view of the above circumstances, and an object thereof is to save energy for operating a torque distribution drive source.

[0006]

In order to achieve the above object, the invention described in claim 1 is a torque distribution drive for distributing and controlling engine torque to at least two output shafts at a predetermined ratio. In a vehicle torque distribution device including a power source, the torque distribution drive source is configured by a hydraulic motor driven by a variable displacement hydraulic pump, and the hydraulic pump is configured to discharge pressure oil during torque distribution control. It is characterized by having done.

In addition to the structure of claim 1, the invention described in claim 2 is a swash plate type hydraulic pump in which the hydraulic pump is driven at a rotational speed according to the vehicle speed. The tilt angle of the swash plate is controlled according to the steering angle.

[0008]

Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows a vehicle torque distribution apparatus according to a first embodiment of the present invention applied to a front engine / front drive vehicle. As shown in the figure, a mission M is connected to an engine E mounted horizontally on the vehicle body, and a differential device input shaft 1 which is an output shaft of the mission M.
Includes an input gear 2 for transmitting a driving force to a planetary gear type main differential device D.

The main differential device D includes a ring gear 4 having an outer tooth gear 3 on its outer periphery that meshes with an input gear 2 of the differential device input shaft 1, and a sun gear 5 coaxially arranged inside the ring gear 4. , An outer planetary gear 6 that meshes with the ring gear 4 and an inner planetary gear 7 that meshes with the sun gear 5, and a planetary carrier 8 that supports the inner planetary gear 7 in a mutually meshing state. In the main differential device D, the ring gear 4 functions as an input element, and the planetary carrier 8 that functions as one output element is connected to the right wheel W _R via the right shaft 9 and functions as the other output element. Sun gear 5 is left shaft 1
It is connected to the left wheel W _L via 0.

Next, the structure of the planetary gear device P which distributes the torque input from the ring gear 4 which is the input element of the main differential device D to the planetary carrier 8 and the sun gear 5 which are the two output elements at a predetermined ratio. Will be explained.

In the planetary gear set P, the left shaft 10
Planetary carrier 12 as first element coupled to
The planetary gear 13 provided in the sun gear 1 as a third element rotatably supported on the left shaft 10
4 and also meshes with a ring gear 15 as a second element arranged on the outer periphery of the planetary carrier 12. The external gear 16 formed integrally with the planetary carrier 8 of the main differential device D and the external gear 17 formed on the ring gear 15 of the planetary gear set P mesh with a pair of integrally formed pinions 18 and 19, respectively. To do. As a result, the main differential device D and the planetary gear device P are connected to each other.

Now, the planetary gear 1 of the planetary gear unit P
3, the numbers of teeth of the sun gear 14 and the ring gear 15 are Z _P , Z _S , and Z _R , respectively, and the rotational speeds of the planetary carrier 12, the sun gear 14, and the ring gear 15 are ω.
_C, ω _S , ω _R , when the sun gear 14 is fixed (that is, ω _S = 0), as is well known, ω _R = ω _C × (1 + Z _S / Z _R ) ... Holds.

Considering the case where the right wheel W _R and the left wheel W _L rotate at the same speed, the rotation speed of the planetary carrier 12 of the planetary gear train P that rotates integrally with the left wheel W _L is as described above. ω _C , and the rotation speed of the planetary carrier 8 of the main differential D that rotates together with the right wheel W _R that is the same speed as the left wheel W _L also becomes ω _C. The rotational speed ω _R of the ring gear 15 driven by the planetary carrier 12 of the planetary gear device P is ω _C ×
It is represented by (1 + Z _S / Z _R ).

That is, in order for the right wheel W _R and the left wheel W _L to rotate at the same speed ω _C , the rotation speed of the planetary carrier 8 of the main differential device D becomes ω _C , and the ring gear 15 of the planetary gear device P becomes. Rotation speed is ω _C × (1 + Z _S / Z _R ).
Therefore, it is necessary to interlock the planetary carrier 8 and the ring gear 15 with each other by the pair of pinions 18 and 19. For that purpose, the radius r ₁ of the external gear 17 formed on the ring gear 15 and the planetary carrier 8
The radius r ₂ of the externally toothed gear 16 formed in ₁ may be set so as to satisfy the relationship of r ₂ / r ₁ = 1 + (Z _S / Z _R ).

A known axial piston type variable displacement hydraulic pump 20 is connected to and driven by a pinion 21 that meshes with an external gear 3 formed integrally with a ring gear 4. The hydraulic pump 20 is connected to a hydraulic motor 24 via oil passages 22 and 23, and a pinion 25 provided on an output shaft of the hydraulic motor 24 has a sun gear 14 of the planetary gear device P.
And a planetary gear device input gear 26 integrally formed with the gear.

Both ends of the steering gear 27 that moves in the left-right direction of the vehicle body in conjunction with the operation of the steering wheel are
It is connected via a pair of Bowden wires 28, 29 to a capacity adjusting lever 30 for driving the swash plate of the hydraulic pump 20. Therefore, the discharge amount of the hydraulic pump 20 becomes 0 when the steering wheel is in the neutral state, and when the steering wheel is steered in one direction, the steering angle and the rotation speed of the pinion 21 (that is, the vehicle speed) at that time are determined. When a certain amount of pressure oil is discharged from the hydraulic pump 20 to the oil passage 22 side and the steering wheel is steered in the other direction, an amount of pressure oil corresponding to the steering angle and the vehicle speed at that time is generated.
Is discharged to the oil passage 23 side.

As is apparent from FIG. 2, the hydraulic control means 34 connected to the electronic control unit U includes a strainer 36, a charge pump 37, a pressure regulating valve 38, and a pair of oil pressure regulators between the oil tank 35 and the oil passages 22 and 23. Check valves 39 and 40, and a relief valve 41 and a pair of check valves 42 and 43 between the pressure regulating valve 38 and the oil passages 22 and 23. The relief valve 41 is connected to an electronic control unit U and controlled by a linear solenoid 44, and a valve spring 45 capable of adjusting a set load by the linear solenoid 44.
And a valve element 46 biased in the valve closing direction by the valve spring 45, and opens when the hydraulic pressure in the oil passages 22 and 23 exceeds a relief pressure determined by the set load of the valve spring 45. To do. When the oil pressure of one of the oil passages 22 and 23 exceeds the relief pressure of the relief valve 41, the oil pressure is released to the other oil passages 22 and 23 via the check valve 39 or the check valve 40. The oil lost from the oil passages 22 and 23 is replenished from the charge pump 37 to the oil passages 22 and 23 via the pressure regulating valve 38 and the check valves 39 and 40.

Oil pressure sensors 47 and 48 connected to the electronic control unit U are provided in the oil passages 22 and 23, respectively, and the relief pressure of the relief valve 41 is adjusted based on the outputs of the oil pressure sensors 47 and 48. By adjusting, the oil pressure of the oil passages 22 and 23 is feedback-controlled, and the rotation speed of the hydraulic motor 24 can be adjusted.

Next, the operation of the first embodiment of the present invention having the above construction will be described.

During normal running of the vehicle, the torque distribution mode is applied, and the command load from the electronic control unit U increases the set load of the valve spring 45 of the relief valve 41 of the hydraulic control means 34. As a result, the oil passages 22 and 23 are prevented from communicating with each other, and the oil discharged from the hydraulic pump 20 is discharged from the hydraulic motor 2.
4 is supplied. Since the discharge direction of the hydraulic pump 20 is determined by the steering direction and the discharge amount is determined by the steering angle and the vehicle speed, the hydraulic motor 24 is driven according to the steering direction and the steering angle of the steering wheel and the vehicle speed. It will be.

When the vehicle is running straight without the steering wheel being operated, the displacement adjusting lever 30 for driving the swash plate of the hydraulic pump 20 is held in the neutral position, so that the discharge amount of the hydraulic pump 20 becomes zero and the hydraulic motor 24 Is held in a stopped state, and the sun gear 14 of the planetary gear device P connected to the pinion 25 of the hydraulic motor 24 via the planetary gear device input gear 26 is fixed. At this time, the planetary carrier 8 of the main differential device D and the planetary carrier 12 of the planetary gear device P are predetermined via the ring gear 15, the external gear 17, the pinion 19, the pinion 18, and the external gear 16 as described above. The gear ratios are linked together. Therefore, the rotational speeds of the planetary carriers 8 and 12, that is, the rotational speeds of the planetary carrier 8 and the sun gear 5, which are a pair of output elements of the main differential device D, are forcibly matched, and the right wheel W _R and the left wheel W _L are matched. And rotate at the same speed.

When the steering wheel is operated to turn the vehicle, the hydraulic pump 20 is driven according to the steering direction, the steering angle and the vehicle speed, and the hydraulic oil is discharged to cause the hydraulic motor 24 to rotate in a predetermined rotational direction and rotational speed. To rotate. As a result, the sun gear 14 of the planetary gear device P rotates, and the rotation speeds of the planetary carriers 8 and 12, that is, the planetary carrier 8 and the sun gear 5 of the main differential device D are increased.
A predetermined difference occurs in the rotation speed of the. Thus, the torque transmitted from the mission M to the ring gear 4 of the main differential device D is transmitted to the left and right wheels W _L , W _R at a predetermined ratio determined by the rotation direction and rotation speed of the hydraulic motor 24. .

At this time, the rotational speed ω ₁ of the right wheel W _R and the left wheel W
The discharge amount of the hydraulic pump 20 is adjusted to determine the rotational speed of the hydraulic motor 24 so that the rotational speed difference Δω between the rotational speed ω _{2 of L and} the rotational speed ω ₂ is appropriate for the turning state of the vehicle.

This will be described with reference to FIG. Wheelbase L, a vehicle of the tread T is, the vehicle speed V, the steering angle theta, the turning radius R of the vehicle, right wheel W _R the turning radius R ₁ of the left wheel W _L turning radius R ₂ of the yaw rate Y (= V / R) in assuming that the left turn, the rotational speed difference [Delta] [omega between the rotational speed omega ₂ of the right wheel W _R rpm omega ₁ and left wheel W _{L of, Δω = ω 1 -ω 2 αY} × R 1 - It is represented by Y × R ₂ = Y × T = (V / R) × T. Therefore, the rotational speed difference Δ between the left and right wheels W _L , W _R
ω is proportional to the vehicle speed V and inversely proportional to the turning radius.

Next, considering the relationship between the turning radius R and the steering angle θ, from FIG. 1, R sin θ = L, and when θ is small, R≈L / θ. , Δω∝ (T / L) × V × θ. The equation shows that the rotational speed difference Δω between the left and right wheels W _L and W _R is proportional to the vehicle speed V and also to the steering angle θ.

The rotational speed difference Δω is determined by the hydraulic motor 24.
Is proportional to the rotation speed of the hydraulic motor 24, and the rotation speed of the hydraulic motor 24 is constant if the capacity Q _M of the hydraulic motor 24 is constant.
It is proportional to the discharge amount Q _P 0. Further, as is well known, when the swash plate tilt angle α is small, the hydraulic pump 20
The discharge amount Q _P is proportional to the swash plate tilt angle α and proportional to the rotation speed of the hydraulic pump 20 (that is, the vehicle speed V). Therefore, the rotation speed difference Δω is represented by Δω∝Q _P ∝α × V.

From the above equations and expressions, α∝ (T / L) × θ ... This expression is obtained by making the swash plate tilt angle α of the hydraulic pump 20 proportional to the steering angle θ when the hydraulic pump 20 is driven in proportion to the vehicle speed V, that is, the steering wheel as in this embodiment. The capacity adjusting lever 3 for driving the swash plate of the hydraulic pump 20 via the Bowden wires 28 and 29 by the steering gear 27 interlocked with the operation of
By connecting to 0, between the left and right wheels W _L , W _R ,
Turning state at that time (that is, vehicle speed V and steering angle θ)
It is shown that it is possible to generate a proper rotation speed difference Δω corresponding to the above condition and perform stable turning. still,
Even if the hydraulic pump 20 is not driven according to the vehicle speed, and is driven at a constant rotation speed, for example, the swash plate tilt angle α of the hydraulic pump 20 is proportional to the product of the vehicle speed V and the steering angle θ. Thus, the same rotation speed difference Δω as described above can be obtained.

Further, by weakening the set load of the valve spring 45 of the relief valve 41 and lowering the relief pressure in response to a command from the electronic control unit U, a part of the oil discharged from the hydraulic pump 20 is discharged between the oil passages 22 and 23. Hydraulic motor 24
It is possible to appropriately adjust the distribution of the torque to the left and right wheels W _L and W _R by reducing the rotational speed of

As described above, the discharge amount of the hydraulic pump 20 is set to 0 when the steering wheel is not operated,
Further, when the steering wheel is operated, the discharge amount of the hydraulic pump 20 increases in proportion to the steering angle. Therefore, the energy required to drive the hydraulic pump 20 is suppressed to the minimum necessary, and the output of the engine E is wasted. It is never consumed.

Next, a second embodiment of the present invention will be described with reference to FIG.

In the second embodiment, the swash plate of the variable displacement hydraulic pump 20 driven according to the vehicle speed is connected to the electric motor 49, and the electronic control unit U from the electronic control unit U connected to the steering angle sensor 31. By driving the electric motor 49 according to the command, the discharge direction and the discharge amount of the hydraulic pump 20 are controlled. The oil passages 22 and 23 are connected to the oil tank 35 via the shuttle valve 50 and the relief valve 51. The other structure is the same as that of the first embodiment.

Then, with the set load of the relief valve 41 being increased to prevent the oil passages 22 and 23 from being short-circuited, the electric motor 49 is operated based on the steering angle detected by the steering angle sensor 31.
Is driven to control the tilt angle of the swash plate of the hydraulic pump 20 so as to be proportional to the steering angle. At this time, since the rotation speed of the hydraulic pump 20 is proportional to the vehicle speed, the discharge amount of the hydraulic pump 20 is eventually proportional to the vehicle speed and the steering angle, and the left and right wheels W _L and W _R are the same as in the first embodiment. It is possible to generate a proper rotation speed difference Δω corresponding to the turning state in the meantime and perform a stable turning.

In the torque distribution mode, the relatively high temperature oil discharged from the hydraulic motor 24 to the oil passages 22 and 23 on the low pressure side is supplied to the shuttle valve 50 and the relief valve 51.
It is released to the oil tank 35 via.

Also in this second embodiment, the discharge amount of the hydraulic pump 20 is set to 0 when the steering wheel is not operated, and when the steering wheel is operated, the discharge amount proportional to the steering angle is generated. The driving energy of the pump 20 can be minimized.

Next, a third embodiment of the present invention will be described with reference to FIG.

In the third embodiment, the driving force of the electric motor 49 in the second embodiment is amplified by the hydraulic assist mechanism 52 to drive the swash plate of the hydraulic pump 20. The hydraulic assist mechanism 52 includes a hydraulic pump 2 inside the casing 53.
The first spool 54 connected to the No. 0 swash plate is slidably accommodated, and the electric motor 4 is placed inside the first spool 54.
The second spool 55 connected to 9 is slidably accommodated, and the high pressure (P ₁ ) first oil chamber 5 connected to the pressure regulating valve 38 is connected.
6, a second oil chamber 58 of medium pressure (P ₂ ) connected to the oil tank 35 via the check valve 57, and a third oil chamber 59 of low pressure (P ₃ ≈0) connected to the oil tank 35.

Thus, when the electric motor 49 is stopped, both the first valve portion V ₁ and the second valve portion V ₂ formed between the first spool 54 and the second spool 55 are closed. And the load by which the hydraulic pressure P ₁ of the first oil chamber 56 pushes the first spool 54 to the right and the load by which the hydraulic pressure P ₂ of the second oil chamber 58 pushes the first spool 54 to the left are balanced. The first spool 54 is stopped. When the second spool 55 is moved to the right by the electric motor 49, the second valve portion V ₂ is opened and the oil pressure in the second oil chamber 58 is reduced from P ₂ to P ₃ , so that the oil pressure P _{1 in} the first oil chamber 56 is reduced. The first spool 54 moves right by being pushed by,
It stops at the position where the second valve portion V ₂ is closed. Conversely, when the second spool 55 is moved leftward by the electric motor 49, the first valve portion V
₁ is opened and the oil pressure in the second oil chamber 58 is increased from P ₂ to P ₁ , so that the first spool 54 is moved leftward by being pressed by the oil pressure P ₁ in the second oil chamber 58, and the first valve portion V ₁ Stops at the closed position.

As described above, the small capacity electric motor 4
By driving the second spool 55 at 9, hydraulic assist force is exerted, and the first spool 54 can be moved following the second spool 55 to drive the swash plate of the hydraulic pump 20. The check valve 57 provided between the second oil chamber 58 and the oil tank 35 is for fail-safe purpose and is for preventing the second oil chamber 58 from being negatively pressured and hydraulically locked. is there.

Next, a fourth embodiment of the present invention will be described with reference to FIG.

In the fourth embodiment, instead of the electric motor 49 and the hydraulic assist mechanism 52 in the third embodiment, the swash plate of the hydraulic pump 20 is driven by the four-way switching valve 60 and the hydraulic cylinder 61. Hydraulic cylinder 61 is cylinder 6
2 is provided with a piston 63 slidably fitted inside and connected to the swash plate of the hydraulic pump 20.
First oil chamber 65 and second oil chamber 66 formed on both sides of
Further, the hydraulic pressure is selectively supplied via the four-way switching valve 60 having the linear solenoids 67 and 68 and the spool 69.

When the spool 69 of the four-way switching valve 60 is moved to the left by a command from the electronic control unit U connected to the steering angle sensor 31, hydraulic pressure is supplied to the first oil chamber 65 of the hydraulic cylinder 61. And the piston 63 moves to the right, again 4
When the spool of the direction switching valve 60 is moved to the right, the hydraulic pressure is supplied to the second oil chamber 66 of the hydraulic cylinder 61 to cause the piston 63 to move.
Moves to the left, whereby the swash plate of the hydraulic pump 20 can be driven in proportion to the steering angle.

Next, a fifth embodiment of the present invention will be described with reference to FIGS. 7 and 8.

As is apparent from FIG. 7, the hydraulic pump 20 of the fifth embodiment is a variable displacement pump having a constant discharge direction, and the displacement adjusting lever 30 biased in one direction by the spring 33 is the steering wheel. The Bowden wire 28, so that the wheel swings in the same direction no matter which direction it is steered,
29 is connected. Therefore, the discharge amount of the hydraulic pump 20 is 0 when the steering wheel is not operated, and no matter which direction the steering wheel is operated, one of the oil passages (oil passage 23) is increased according to the increase of the steering angle.
The discharge amount to

As shown in FIG. 8, a relief valve 70 having the same structure as the relief valve 41 and a check valve 71 are provided between the oil passage 23 on the discharge side and the oil passage 22 on the return side of the hydraulic pump 20. Is provided. The oil passage 23 is bifurcated, one of which is connected to one port of the hydraulic motor 24 via a check valve 72 and a pressure regulating valve 73 having a linear solenoid, and the other of which is connected to a check valve 74 and a linear solenoid. It is connected to the other port of the hydraulic motor 24 via a pressure regulating valve 75 provided. The oil passage 22 is also bifurcated and connected to the pair of pressure regulating valves 73 and 75, respectively.

When the torque distribution mode is selected,
Steering direction sensor 3 connected to the electronic control unit U
2 For example, the steering wheel is steered to the right.
If it is detected that the pressure regulating valve 73, 75 is
By switching each to the left position and the right position, the hydraulic pressure
The motor 24 is driven in one direction and, conversely, the steering wheel
If it is detected that the steering wheel has been steered to the left, a pair of
Turn pressure regulators 73 and 75 to the right and left positions respectively.
The hydraulic motor 24 is driven in the other direction
It At this time, the rotation speed of the hydraulic pump 20 is proportional to the vehicle speed.
And the swash plate tilt angle of the hydraulic pump is proportional to the steering angle.
Therefore, the discharge amount of the hydraulic pump, that is, the hydraulic motor
The rotation speed of the motor 24 is proportional to the vehicle speed and the steering angle.
Left and right wheels W _L, W_RThe rotation speed difference Δω is commensurate with the turning state
It will be appropriate.

In the torque distribution mode, the relief pressure of the relief valve 70 is increased to supply the entire amount of oil discharged from the hydraulic pump 20 to the hydraulic motor 24.
Relief pressure of relief valve 70 is reduced to reduce hydraulic pump 2
If a part of the discharged oil of 0 is recirculated without passing through the hydraulic motor 24, the differential force generated in the main differential device D can be adjusted appropriately.

When the diff lock mode is selected,
Both the pair of pressure regulating valves 73 and 75 are switched to the left position, and the relief valve 70 is substantially opened. As a result, the oil discharged from the hydraulic pump 20 flows back without load via the relief valve 70, and the hydraulic motor 24 is locked by the pair of check valves 72 and 74 so as not to rotate. In this diff lock mode, the rotational speeds of the left and right wheels W _L and W _R are forcibly matched to realize a so-called diff lock state, which is effective when the vehicle is slipped into a mud and slips into one wheel. To do.

Further, when the normal differential mode is selected, both the pair of pressure regulating valves 73 and 75 are switched to the right position, and the relief valve 70 is substantially opened.
As a result, the oil discharged from the hydraulic pump 20 flows back without load via the relief valve 70, and the hydraulic motor 24 is connected in a closed circuit via the pair of pressure regulating valves 73 and 75 so that it can freely rotate without load. Then, the normal differential mode is realized.

In this normal diff mode, if the pair of pressure regulating valves 73 and 75 are both urged toward the left position by the linear solenoids, an intermediate state between the normal diff mode and the diff lock mode, that is, the main differential device is obtained. It is possible to realize a state in which a predetermined operation limiting force acts on D.

Although the embodiments of the present invention have been described above in detail, the present invention is not limited to the above embodiments, and various design changes can be made.

For example, the configuration for torque distribution is not limited to that of the embodiment using the main differential device D and the planetary gear device P, and various known devices can be substituted.
Further, the hydraulic pump 20 is not limited to the swash plate type axial piston pump, and a swash plate type radial piston pump or a vane pump capable of adjusting the capacity by changing the eccentric amount of the cam ring can be employed. Further, when the rotation speed of the hydraulic pump 20 is not proportional to the vehicle speed, for example, when the hydraulic pump 20 is directly connected to the crankshaft of the engine E to be driven or is driven by an electric motor, the discharge amount of the hydraulic pump 20 is set to the vehicle speed and steering angle. By controlling so as to be proportional to both, it is possible to obtain an appropriate rotational speed difference Δω between the left and right wheels W _L and W _R. At this time, the discharge amount from the hydraulic pump 20 may be set to be slightly larger than the required amount, and the surplus amount may be discharged from the relief valves 41 and 70. It is also possible to discharge a small amount of oil from the hydraulic pump 20 even when the torque distribution control is not being performed. Further, the torque distribution mechanism of the present invention is applicable not only to the drive system for the front wheels of the vehicle but also to the drive system for the rear wheels, and also applicable to the torque distribution between the front wheels and the rear wheels in a four-wheel drive vehicle.

[0053]

As described above, according to the invention described in claim 1, the torque distribution drive source is constituted by a hydraulic motor driven by a variable displacement hydraulic pump, and the hydraulic pump is torque distributed. Since the pressure oil is discharged during control, it is possible to avoid waste of the hydraulic pump discharging pressure oil when torque distribution is not performed, and adjust the discharge amount of the hydraulic pump during torque distribution to achieve the required amount. It is possible to discharge only the pressure oil of and to minimize the energy for driving the hydraulic pump.

According to the second aspect of the invention, the hydraulic pump is a swash plate type hydraulic pump driven at a rotational speed according to the vehicle speed, and the swash plate tilt angle of the swash plate type hydraulic pump is steered. By controlling according to the angle, the rotation speed of the hydraulic motor, that is, the ratio of torque distribution can be made appropriate according to the turning state of the vehicle.

[Brief description of drawings]

FIG. 1 is a diagram showing a torque distribution device according to a first embodiment.

FIG. 2 is a diagram showing a hydraulic circuit of the first embodiment.

FIG. 3 is an explanatory diagram of operation

FIG. 4 is a diagram showing a hydraulic circuit of a second embodiment.

FIG. 5 is a diagram showing a hydraulic circuit of a third embodiment.

FIG. 6 is a diagram showing a hydraulic circuit of a fourth embodiment.

FIG. 7 is a diagram showing a torque distribution device according to a fifth embodiment.

FIG. 8 is a diagram showing a hydraulic circuit of a fifth embodiment.

[Explanation of symbols]

 9 Right Shaft (Output Shaft) 10 Left Shaft (Output Shaft) 20 Hydraulic Pump 24 Hydraulic Motor (Drive Source for Torque Distribution) E Engine

Claims (2)

[Claims] 1. Engine (E) torque of at least 2
A torque distribution device for a vehicle, comprising a torque distribution drive source (24) for distribution control to one output shaft (9, 10) at a predetermined ratio, wherein the torque distribution drive source (24) is a variable displacement type. It consists of a hydraulic motor driven by a hydraulic pump (20),
A torque distribution device for a vehicle, wherein the hydraulic pump (20) is configured to discharge pressure oil during torque distribution control. 2. The hydraulic pump (20) is a swash plate hydraulic pump driven at a rotational speed according to the vehicle speed, and the swash plate tilt angle of the swash plate hydraulic pump is controlled according to the steering angle. The torque distribution device for a vehicle according to claim 1, wherein: