GB2280157A - Torque distribution control system for a four-wheel-drive motor vehicle - Google Patents

Abstract

The vehicle has a central differential to control the torque distribution between the front and rear wheels and a rear differential to control the torque distribution between the rear wheels on either side. The restricting torque for the differentials is controlled in accordance with the sensed parameters of; engine speed (45), accelerator position (46), gear selection (47), longitudinal acceleration (44) and lateral acceleration (41). The system calculates and engine torque, desired steering characteristic and road/wheel friction coefficient from the parameters in order to determine an optimum torque distribution to be provided by each of the central and rear differentials (20, 11) at all driving conditions in order to stop vehicle slip. <IMAGE>

Description

Torque Distribution Control System for a Four-Wheel Drive Motor Vehicle The present invention relates to a torque distribution control system for a four-wheel drive motor vehicle having a central differential, and more particularly to the control system in which output torque of a transmission is unequally distributed to front wheels and rear wheels in accordance with driving conditions of the motor vehicle. A differential operation restricting clutch is provided in the central differential for restricting the differential operation. The torque distributed to the front wheels and rear wheels is controlled by controlling the restricting torque of the clutch.

It is known that driving performance of the motor vehicle differs with the type of the power transmission system. For example, in the four-wheel drive motor vehicle, the four wheels are driven to prevent slipping and skid of the wheels, thereby improving driving performance in traction, braking, and steering. The acceleration or deceleration influences on the front wheels and the rear wheels at the same time, so that both the understeer and oversteer of the vehicle are reduced.

Furthermore, in the four-wheel drive motor vehicle, the torque distribution to front wheels and rear wheels and to left rear-wheel and right rear-wheel affects on the steerability and the change of running behavior. If the torque distribution is properly controlled, driveability and dynamic stability are further improved. Consequently, it has been proposed to properly and variably control the torque distribution in dependency on the driving conditions.

Japanese Patent Application Laid-open 63-13824 discloses a torque distribution control system having a central differential for a four-wheel drive motor vehicle. In the system, a fluid operated multiple-disk friction clutch is provided in the central differential for restricting the differential operation. The torque distribution to the front and rear wheels is controlled by controlling the restricting torque. The cornering behavior of the vehicle can be detected by lateral acceleration. If the lateral acceleration increases, the gripping force of the tire gradually reduces to a breakaway point, at which the tires begin to slide sideways to cause spinning or drifting of the vehicle.

Therefore, the restricting torque is set in dependency on the lateral acceleration for controlling the torque distribution to front and rear wheels, thereby preventing the sliding of the tires.

However, the cornering condition can be detected only in a linear zone where the side force changes linearly.

When the gripping force of the tire approaches that of the breakaway point, the side force changes irregularly.

Consequently, the actual lateral acceleration changes irregularly with the behaviour or the vehicle. Therefore, the system can not prevent the vehicle from spinning and drifting in the non-linear side force zone.

An object of the present invention is to provide a torque distribution control system for a four-wheel drive motor vehicle which may ensure driveability, driving stability and steerability even if a gripping force of tires approaches that of a breakaway point.

According to the present invention there is provided a control system for distributing torque to each wheel of a four-wheel drive motor vehicle having: a central differential acting between the front and rear wheels to absorb any speed difference between the front and rear wheels, a central clutch mounted on said central differential to control the distribution of torque to the wheels, a rear differential acting to distribute torque between the rear wheels and having a rear clutch arranged to control the torque distribution of the rear differential, a vehicle speed sensor, a lateral acceleration sensor for detecting lateral acceleration, a steering angle sensor, an engine speed sensor, an accelerator position sensor, a gear position sensor, a longitudinal acceleration sensor sensitive to any longitudinal acceleration of the vehicle, input torque determining means responsive to the engine speed, the accelerator position and the gear position to determine the input torque, steering calculator means responsive to the longitudinal acceleration to calculate a desired steering characteristic, friction coefficient setting means for setting a friction coefficient in response to the lateral and longitudinal acceleration, torque distribution calculating means responsive to the desired steering characteristic, the longitudinal acceleration and the friction coefficient to calculate a desired torque distribution ratio, a torque calculator means responsive to the input torque, and the torque distribution ratio to calculate a first control value for controling the central clutch to minimise slipping of the motor vehicle and to stabilise the vehicle under braking, and a rear torque calculator responsive to the steering characteristic, the longitudinal acceleration and the friction coefficient to calculate a rear clutch control value to control the rear clutch so as to control the rear differential torque distribution to minimise the degree of spin of the vehicle and to stabilise the attitude of the vehicle.

A control system for distributing torque to each wheel of a four wheel drive motor vehicle embodying the present invention will now be described, by way of example only, with reference to the accompanying drawings, in which Fig. 1 is a schematic illustration, Fig. 2 is a block diagram, Fig. 3 is a steering characteristic of stability factor against longitudinal acceleration, Fig. 4 is a controlling characteristics of a torque distribution ratio for distributing the torque to the front wheels and the rear wheels; Fig. 5 is a graph showing'the controlling characteristics of a rear differential restricting torque; Fig. 6 is a power transmission system showing a second embodiment; Fig. 7 is a block diagram of the second

Fig. 1 shows a power transmission system for a four-wheel drive motor vehicle having a central differential according to the present invention. An engine 1 is mounted in a front portion of the vehicle.

A clutch 2 and a transmission 3 are disposed at a rear of the engine 1 in a longitudinal direction of the vehicle. An output of the transmission 3 is transmitted to an output shaft 4 which is aligned with a central differential 20. The output shaft 4 is connected to a front drive shaft 5 which isparallelly disposed under the transmission 3 through a pair of reduction gears 25 and 26 of the central differential 20. The front drive shaft 5 is connected to left and right front wheels 9L and 9 through a front differential 7 and axles 8. The output shaft 4 is connected to a rear drive shaft 6 through the central differential 20. The rear drive shaft 6 is connected to the left and right rear wheels 13L and 13R through a propeller shaft 10, a rear differential 11 and axles 12.

The central differential 20 is a complex planetary gear device and comprises a first sun gear 21 integrally formed on the output shaft 4, a second sun gear 22 integrally formed on the rear drive shaft 6, and a combined planetary pinion 23 comprising a first planetary pinion 23a meshed with the first sun gear 21, and a second planetary pinion 23b meshed with the second sun gear 22, and supported on a carrier 24. The carrier 24 is connected to the reduction drive gear 25.

Thus, an output torque from the output shaft 4 of the transmission 3 is transmitted to the carrier 24 and the second sun gear 22 through the first sun gear 21 and the pinions 23a, 23b at predetermined respective torque distribution ratios. A difference between rotating speeds of the carrier 24 and the second sun gear 22 is absorbeby.rot.ation of the first and second planetary pinions-23a and 23b.

Consequently, a standard torque distribution for front torque TF and rear torque TR can be set to various values by changing radii of pitch circles of the sun gears 21 and 22 and the pinions 23a and 23b.

Thus, the torque distribution ratio et of the front wheels 9L, 9R and the rear wheels 13L, 13R is determined as follows: Tp : TR . 34 : 66 A large standard torque can be distributed to the rear wheels 13L, 13R.

A fluid operated multiple-disk friction clutch 27 is'provided adjacent the central differential 20 for restricting the differential operation of the central differential 20.

The clutch 27 comprises a drive drum 27a secured to the carrier 24, and a driven drum 27b secured to the rear drive shaft 6. When a differential operation restricting clutch torque Tc is produced in the clutch 27, a part of the output torque of the second sun gear 22 is transmitted to the front wheel 9L, 9R, thereby changing the distribution of the torque. The carrier 24 is coupled with the second sun gear 22 when the clutch 27 is entirely engaged, thereby locking the central differential 20.

In the vehicle with the front-mounted engine, static weight distribution ew of front dynamic weight WF and rear dynamic weight WR are determined as follows; WF : WR =. 62 : 38 When the clutch 27 is directly engaged, the distribution ratio et of front torque and rear torque is set in accordance with the weight distribution ew.

Thus, the torque distribution is controlled in a range between the standard torque distribution of 34:66 weighted to the rear wheels 13L, 13R and a torque distribution of 62:38, weighted to the front wheels 9L, 9R at an entire engagement of the clutch 27 in accordance with the differential operation restricting clutch torque Tc.

A hydraulic circuit for controlling the clutch 27 will be described hereinafter.

The hydraulic circuit having a control system 32 for the clutch 27 comprises an oil pump 30, a pressure regulator valve 31, a pilot valve 36, an clutch control valve 34 and a solenoid operated duty control valve 40.

The regulator valve 31 operates to regulate a pressure of oil supplied from the oil pump 30 driven by the engine 1 to produce a line pressure and the line pressure is applied to a line pressure conduit 33. The conduit 33 is communicated with a passage 38 through the pilot valve 36. The passage 38 is communicated with the-solenoid operated duty control valve 40 at downstream of an orifice 37, and with an end of the clutch control valve 34 through a passage 39. The conduit 33 is communicated with the clutch control valve 34 through a passage 33a. The clutch control valve 34 is communicated with the clutch 27 through a passage 35. The solenoid operated valve 40 is oprated by pulses from a control unit 50 at a duty ratio determined therein, to control drainage of the oil to provide a control pressure Pd. The control pressure Pd is applied to the clutch control valve 34 to control the oil supplied to the clutch 27 so as to control the clutch pressure (torque) and hence the differential operation restricting torque Tc.

A power transmission system is provided with a fluid operated multiple-disk friction clutch 28 provided adjacent the rear differential 11 for restricting the differential operation of the differential 11. The rear differential 11 comprises a bevel gear differential device ounted in a differential case lia. The clutch 28 comprises a drive drum 28a secured to the differential case lla, and a driven drum 28b secured to one of the axles 12 connected to a side gear llb of the differential 11.

When the clutch 28 is released, the torques are equally distributed to the left and right wheels 13L and 13R. When the clutch 28 is engaged, thereby generating a differential operation restricting torque Td, the differential operation of the rear differential 11 is restricted. The torque distribution ratio of the left and right wheels are determined in accordance with left and right dynamic weights W on the left and right wheels 13L and 13R, and the friction coefficient y of the road surface (W.i).

A hydraulic circuit for controlling the clutch 28 will be described hereinafter. A control system 32' for the clutch 28 is operatively connected to the control system 32 for the clutch 27.

The control system 32' for che clutch 28 comprises a clutch control valve 34' and a solenoid operated duty control valve 40'. The passage 38 from the pilot valve 36 is communicated with a passage 38' which is communicated with the solenoid operated duty control valve 40' at downstream of an orifice 37', and with an end of the clutch control valve 34' through a passage 39'. The conduit 33 is communicated with the clutch control valve 34' through a passage 33b. The clutch control valve 34' is communicated with the clutch 28 through a passage 35'. The solenoid operated valve 40' is operated in the same manner as the solenoid operated valve 40 to provide a control pressure Pd'. The control pressure is applied to the clutch control valve 40' to control the oil supplied to the clutch 28 so as to. control the clutch pressure (torque) and hence the differential operation restricting torque Td.

Fig. 2shows the control unit 50 A-ystem for controlling the torque distribution to the front wheels and the rear wheels and to a left rear-wheel and a right rear-wheel will be described.

The principle of the control system is described first. In order to obtain stable driving of the vehicle during the driving on the dry road, and to prevent the spin of the vehicle for obtaining stability on the slippery roadr it is necessary to control the torque distribution so as to provide a good steering characteristic in various road conditions and driving conditions, namely to control the stability factor to be constant.

If the vehicle makes a turn at acceleration during the four-wheel driving, the cornering power and the side force at the front tires are reduced due to longitudinal weight transfer, and the cornering power and the side force at the rear tires is increased, causing to an understeer. The relationship between the driving power and the side force of the tire is determined in accordance with a friction circle (i.e. a circle of static function force) in accordance with the friction coefficient z of the road and the load on the vehicle. If the driving power is increased, the side force reduces, which affects the steering characteristic. Therefore, in order to obtain a desired steering characteristics, the driving power of the rear tires is increased to reduce the side force for cancelling the increase of the cornering power of the rear tires.

In the rear tires driving, a yawing moment is produced on the vehicle body, which influences the steering characteristic. The driving conditions of the rear tires depend on the input driving power. If the input driving power is small or the engine braking is effected, the outside tire of the rear tires is braked and the driving power of the inside tire increases, so that the yawing moment is produced in an understeer direction. If the input driving power is large, the driving power produces in dependency on the contact load of the tires, so that the yawing moment is produced in an oversteer direction; Therefore, the torque distribution to the left rear tire and the right rear tire is controlled so as to maintain a good steering characteristic in accordance with the driving conditions.

For an actual controlling of the central differential at acceleration, desired steering characteristics are set at a side with respect to a neutral point, at which the understeer is reduced. To the contrary, for deceleration, the desired steering characteristics are set at a side where the understeer is increased.

Furthermore, the desired steering characteristic is expanded up to the non-linear side force zone and determined in accordance with the stability factor, thereby setting an equivalent cornering power. The cornering power of the front wheels and the rear wheels is obtained by the friction coefficient IL of the road, torque distribution ratio to front and rear wheels, and longitudinal and lateral accelerations. The friction coefficient is substituted with a supposed friction coefficient calculated by square root of a sum which is obtained by adding a square number of the longitudinal acceleration and a square number of the lateral acceleration. The supposed coefficient is smaller than the actual coefficient so that the controlling is performed at a stable side. The supposed coefficient approaches the actual coefficient in the breakaway zone.

It has been found that the cornering power of the tire can be substituted for an equivalent cornering power which is expressed in a predetermined equation by an equation of motion, if the motion at acceleration and deceleration during the cornering, which generates large longitudinal and lateral accelerations, is formalized, extended in the non-linear zone. Thus, if the torque distribution ratio to the front wheels and the rear wheels changes, the equivalent cornering power on the front wheels and the rear wheels is changed, accordingly, as a result, the steering characteristics of the vehicle changes. Therefore, the equivalent cornering power in consideration of non-linearity is a function of longitudinal acceleration G, lateral acceleration, and a supposed friction coefficient s.

Consequently, a torque distribution ratio a to front and rear wheels controlled by the central differential can be calculated by equations as follows.

hw LrL2 A- (Kro KrcGx) 2L 2 3 Lf2 (2) #Yf #Yr #Yf #Yr g C-Lf -Lr-+AlL2 .-.- (3) # f # r # f # r W #Yf hW Gy -2(Kfo-# KrcGx-Kfo) (4) eAf 2L 2u #Yr hW Gy - (Kfo-#KrcGx-Kro-) (5) ##r 2L where /Gx is the longitudinal acceleration, Gy is the lateral acceleration, # is the supposed friction coefficient,A is the desired stability factor, W is the weight of the vehicle, h is the height of the center of gravity, L is the wheelbase, Lf and Lr are distances between the center of gravity and respective axles, Kfo and Kro are equivalent cornering power in a linear zone, and Kfc and Krc are partial differentials of Kf and Kr with respect to W, which represent a load dependent characteristic of the cornering power.

Describing a concrete control of the rear differential, the moment in the understeer direction is proportional to the yawing rate and inversely proportional to the vehicle speed. The moment in the direction of the oversteer is proportional to the driving power to the rear wheels and the lateral acceleration G, since the driving power is large in such a state. Both of the moments are added to calculate the entire yawing moment. The calculated yawing moment is substituted for an equation of motion at a steady circular cornering which is expanded to a non-linear zone and expressed by the stability factor and rewritten in the rear differential restricting torque Td which is calculated by an equation as follows.

where d is the tread, and Rt is the effective diameter of the tire. It is assumed that the equation is plus when Gx 2 0 and minus when Gx < 0.

Referring to Fig. 2, the control unit 50 is provided with a longitudinal G sensor 44 for detecting an acceleration Gx in the longitudinal direction of the vehicle and the lateral G sensor 41. In order to presume an input torque to the central differential 20, there are provided an engine speed sensor 45, an accelerator pedal depression degree sensor 46, and a gear position sensor 47. Further, a signal of an antilock brake system (ABS) control unit 48 is applied to the control unit 50.

The control unit 50 has a friction coefficient setting section 59 supplied with the lonCtudinal acceleration Gx and the lateral acceleration Gy. In the section 59, a friction coefficient sr is calculated from the square root of the sum of the squares of the longitudinal and lateral accelerations

as described above. A desired steering characteristic setting section 60 is supplied with the longitudinal acceleration Gx. In the section 60, a. stability factor is derived from a look-up table as shown in Fig. 7. If the longitudinal acceleration Gx is large at acceleration, a small stability factor Al is determined with respect to the neutral point. If the longitudinal acceleration Gx is small at deceleration, a large stability factor A2 is determined.

An input torque section 61 is supplied with an engine speed N, an accelerator pedal depression degree 3 and a gear position P from the sensors 45, 46 and 47. In the section 61, engine output torque Te is deduced in accordance with the engine speed N and the accelerator pedal depression degree 9 with reference to the output characteristic of the engine. The engine output Te is multiplied by a gear ratio g of the gear position P to calculate input torque Ti.

The lateral and longitudinal accelerations Gy Gx, a stability factor A at acceleration or deceleration, and the friction coefficient sll are applied to a desired torque distribution ratio calculator 62. The torque distribution ratio a to the front wheels and the rear wheels is calculated in accordance with the equations (1) to (5). The torque distribution ratio a and the input torque Ti are applied to a central differential operation restricting torque calculator 63 for calculating the restricting torque Tc for restricting the differential operation of the central differential 20. The torque distribution ratio a is determined between zero for the rear wheels and one for the front-wheels (RWD 0 and FWD 1). If a standard torque distribution ratio Di to the rear wheels is larger than that to the front wheels, the restricting torque Tc is calculated by an equation as follows.

Tc = ( a + Di ) Ti If the calculated value is negative, the torque Tc is set to zero.

The restricting torque Tc is applied to a duty ratio converting section 56 where the torque Tc is converted to a corresponding duty ratio D. The duty ratio D provided at the section 56 is applied to the solenoid operated duty control valve 40.

The lateral and longitudinal accelerations Gx, Gy, stability factor A and the friction coefficient sz are further applied to a rear differential operation restricting torque calculator 57 in which restricting torque Td for restricting the differential operation of the rear differential 11 is calculated in accordance with the equation (6).

The restricting torque Td is applied to a duty ratio converting section 58 where the torque Td is converted to a corresponding duty ratio D. The duty ratio D provided at the section 58 is applied to the solenoid operated duty control valve 40'.

If an ABS control signal is fed from the unit 48 to the calculators 63 and 57, each of the duty ratios D is forcibly corrected so as to set the respective restricting torques Tc and Td to zero.

The operation of the system will be described hereinafter.

In the system, the rear torque determined in accordance with the central differential 20 and the clutch 27 is transmitted to the rear wheels 13L, 13R through the rear differential 11 and the rear clutch 28.

If the clutch 28 is disengaged and the clutch torque becomes zero so as to render the rear differential 11 free. Accordingly, the torque is equally transmitted to the left rear-wheel 13L and the right rear-wheel 13R.

When the restricting torque Td is produced in the rear restricting clutch 28 by the hydraulic control system 32', the rear clutch 28 is engaged and the differential operation of the rear differential 11 is restricted. Thus, the torque is effectively distributed to the rear wheels which grip the ground.

The torque is transmitted from a higher speed wheel to a lower speed wheel responsive to the torque Td. When the rear differential 11 is directly engaged, the torque is distributed to the left rear wheel 13L and the right rear wheel 13R in accordance with weight distribution on the rear wheels.

During the four-wheel driving, in the control unit 50, the supposed friction coefficient SIL is determined in accordance with the lateral acceleration Gy and longitudinal acceleration Gx and the stability factor A is determined in accordance with the longitudinal acceleration Gx for calculating the torque distribution ratio a. In accordance with the torque distribution ratio a and the input torque Ti, the restricting torque Tc is calculated. A duty signal corresponding to the calculated restricting torque Tc is applied to the control system 32 so that the clutch 27 is controlled by feedfoward control to produce the torque Tc.

Fig. 4 shows controlling characteristics of the torque distribution to the front wheels and the rear wheels. At acceleration, the longitudinal acceleration Gx is positive. In the straight-ahead driving where the lateral E mceleration Gy is small, the torque distributed to the front wheels is increased so that the vehicle is driven as a front-drive vehicle. If the vehicle is accelerated on the dry road, the longitudinal acceleration Gx is large so that the torque is equally distributed to the front wheels and the rear wheels, thereby effectively providing running performance and stable driving.

f the vehicle is accelerated at cornering, the lateral acceleration Gy rises. The torgue distributed to the rear wheels is increase, so that the vehicle is driven as a rear-drive vehicle. Thus, the understeer is prevented from increasing td ensure the steering characteristic to be constant. If the vehicle is driven on the slippery road at a small longitudinal acceleration Gx, the torque is equally distributed or the torque to the front wheels is increased, thereby preventing the slipping of the rear wheels.

At deceleration where the longitudinal acceleration Gx is negative, the torque distribution ratio is determined at all times to increase the torque to the front wheels or to directly engage the differential. Thus, the engine braking effect is effected to reduce tuck-in at deceleration.

On the other hand, the restricting torque Td is calculated in the control unit 50 in accordance with the longitudinal and lateral accelerations Gx, Gy, the supposed friction coefficient sp and the stability factor A for controlling the clutch 28 in feedfoward control.

Fig. 5 shows controlling characteristics of the torque distribution to the left rear wheel and the right rear wheel. At acceleration, if the longitudinal acceleration Gy becomes large, a large torque Td is set, so that a moment is produced to reduce the under steer. In this state, the inside wheel of the rear wheels is prevented from slipping to improve the traction. When the vehicle makes a turn at acceleration, the longitudinal and lateral accelerations Gx, Gy becomes large. A large torque is set to lock the differential. When the vehicle is driven on the slippery- road, the torque is set to a minimum so that the differential becomes free, thereby preventing the vehicle from spinning caused.

by the slipping of the rear wheels at the same time.

At deceleration, the torque is increased in dependency on the deceleration speed and the increase of the lateral acceleration Gy. A moment is produced to cause the understeer, thereby preventing tuck-in.

In the first embodiment, the torque distribution ratio is calculated in accordance with parameters of the longitudinal and lateral accelerations Gx, Gy, the friction coefficient sp and the desired steering characteristics by the equation of motion expanded to the non-linear zone of the driving performance.

Therefore, good steerability is obtained in any surface conditions of the road and the driving conditions, thereby improving the steerability at acceleration or deceleration, stable driving, and stability on the slippery road. The rear differential restricting torque is calculated by those parameters so that the same effects as the central differential are obtained.

In particular, the slipping of the wheels, and the spinning and tucking-in of the vehicle are effectively prevented. The supposed friction coefficient sH is used to respond to a rapid change of the surface condition of the road. Since the supposed friction coefficient is smaller than the actual value, safety control is ensured. The supposed coefficient is calculated in accordance with the parameters of the longitudinal and lateral accelerations Gx, Gy and the clutch is controlled in feedfoward control, so that a good response is provided. The desired steering characteristics is determined based on the longitudinal acceleratiqn Gx, so that good steerability at acceleration and stability at deceleration are obtained.

Figs. 7 and 8 show second embodiment of the present invention adapted to an electro-magnetic operated multiple-disk friction clutch i

Claims (5)

1. A control system for distributing torque to each wheel of a four-wheel drive motor vehicle having: a central differential acting between the front and rear wheels to absorb any speed difference between the front and rear wheels, a central clutch mounted on said central differential to control the distribution of torque to the wheels, a rear differential acting to distribute torque between the rear wheels and having a rear clutch arranged to control the torque distribution of the rear differential, a vehicle speed sensor, a lateral acceleration sensor for detecting lateral aceeleration, a steering angle sensor, an engine speed sensor, an accelerator position sensor, a gear position sensor, a longitudinal acceleration sensor sensitive to any longitudinal acceleration of the vehicle, input torque determining means responsive to the engine speed, the accelerator position and the gear position to determine the input torque, steering calculator means responsive to the longitudinal acceleration to calculate a desired steering characteristic, friction coefficient setting means for setting a friction coefficient in response to the lateral and longitudinal acceleration, torque distribution calculating means responsive to the desired steering characteristic, the longitudinal acceleration and the friction coefficient to calculate a desired torque distribution ratio, a torque calculator means responsive to the input torque, and the torque distribution ratio to calculate a first control value for controling the central clutch to minimise slipping of the motor vehicle and to stabilise the vehicle under braking, and a rear torque calculator responsive to the steering characteristic, the longitudinal acceleration and the friction coefficient to calculate a rear clutch control value to control the rear clutch so as to control the rear differential torque distribution to minimise the degree of spin of the vehicle and to stabilise the attitude of the vehicle.

2. A system according to claim 1 wherein each of the central differential restricting torque calculator and the rear differential restricting torque calculator are responsive to an ABS signal from an ABS control unit to set the restricting torques to zero.

3. A system according to claim 1 or claim 2 wherein the clutches are each hydraulically actuated multiple disk friction clutch and each restricting torque calculator controls a duty ratio converting section to drive a solenoid operated valve of the hydraulic control circuit for each clutch.

4. A system according to claim 1 or claim 2 wherein each clutch is an electromagnetically operated multiple disk friction clutch and each restricting torque calculator controls a clutch current converting section to supply a corresponding clutch current to each clutch coil.

5. A control system for distributing torque as herein described with reference to the accompanying drawings.