GB2323940A

GB2323940A - Controlling torque distribution between the rear wheels of a vehicle

Info

Publication number: GB2323940A
Application number: GB9814384A
Authority: GB
Inventors: Koji Matsuno
Original assignee: Fuji Jukogyo KK; Fuji Heavy Industries Ltd
Current assignee: Subaru Corp
Priority date: 1994-06-27
Filing date: 1995-06-26
Publication date: 1998-10-07
Anticipated expiration: 2015-06-26
Also published as: GB9814384D0; GB2323940B

Abstract

The torque distribution between the rear wheels of the vehicle is controlled by calculation of a vehicle yaw moment from the sensed yaw rate, steering angle and vehicle speed.

Description

VEHICLE DRIVING TORQUE DISTRIBUTION CONTROL SYSTEM AND METHOD.

The present invention relates to a control system for controlling the driving torque distribution between front and rear wheels of a vehicle or between a rear-left and rear-right wheel. More specifically it relates to a control system for controlling driving torque distribution based on a calculation using vehicule parameters such as engine output torque, vehicle speed, steering angle, yaw rate and others.

It is well known that a vehicle exhibits a running behaviour unique to its driving system, such as front engine-rear drive system and front engine-front drive system. It is known that a permanent four wheel drive vehicle equipped with a centre differential can has improved marginal performance when the vehicle brakes are applied abruptly or when it turns a corner, compared to FR or F vehicles. Recently, four wheel drive vehicles of this type are becoming popular since they have a medium steering characteristic between over-steer and under-steer and therefore it is said that the four wheel drive vehicles with centre differential are easily operated.

As an example of the torque distribution control system between front and rear wheels of a four-wheel drive vehicle having a centre differential, there is a Japanese Patent Application No. A-63-13824 in which a lateral acceleration during cornering is detected and a differential limiting torque is generated according to the magnitude of the lateral acceleration in a hydraulic multi-disc clutch, thereby the torque distribution between front and rear wheels is controlled so as avoid a spin or drift-out during cornering.

Other examples of a similar technique are found in Japanese Patent Applications A-61-229616 and A-3-74221. The former patent discloses a technique in which the torque distribution between front and rear wheels or between left and right wheels is controlled by detecting spin or driftout based on the difference between a target yaw rate calculated from steering angle and vehicle speed and an actual yaw rate. The latter patent discloses a technique in which the torque distribution between front and rear wheels or between left and right wheels is controlled by calculating a variation versus time of the steering characteristic from steering angle, vehicle speed and actual yaw rate.

However, in the abovementioned prior art techniques, for example in JP-A-63-13824, since the state of turning is detected only by lateral acceleration, the controllable range is limited to a so-called linear grip region where lateral force varies proportionally with respect to a tire slip angle. That is to say, when the vehicle runs on a road with a low friction coefficient and enters into a marginal state such that the grip force of the tire reaches a limit and the vehicle starts to spin, the lateral force varies in a non-linear manner and the actual lateral acceleration varies arbitrarily according to the vehicle behaviour in a spinning state. Consequently, a state of turning of the vehicle cannot be accurately judged. Further, in the prior art disclosed in JP-A-61-229616 and JP-A-3-74221, since the state of turning of the vehicle is judged by a signal from a yaw rate sensor, it is expected that the vehicle behaviour can be judged more accurately than that of JP-A-63-1382a, however, the prior art systems are still insufficient in control in the marginal state.

The present invention is intended to obviate the aforementioned defects and disadvantages of the prior art and it is summarised as follows.

It is an object of the present invention to provide a driving torque distribution control system capable of properly controlling the torque distribution among wheels of a vehicle so as to propel the vehicle with good stability under any running conditions ranging from straight ahead running to cornering and under any road conditions ranging from a road with high friction coefficient to a road with low friction coefficient.

To alleviate the aforementioned disadvantages of the prior art the present invention provides a torque distribution control system for distributing an torque to the wheels of the vehicle, the vehicle having an engine, a transmission connected to the engine via a clutch, a rear differential interposed between a left rear wheel and a right rear wheel to absorb a left and right rear wheel speed difference, a rear solenoid valve hydraulically connected to the rear differential, the system comprising: a steering sensor mounted on the vehicle for sensing a steering angle of the front wheels and for generating a steering angle signal, a vehicle speed sensor mounted on the vehicle for detecting vehicle speed and for producing a vehicle speed signal, a yaw rate sensor mounted on the vehicle for detecting a yaw rate of the vehicle and for generating an actual yaw rate signal: a target yaw rate determining means responsive to the steering angle signal and the vehicle speed signal for setting a target yaw rate based upon a standard turning characteristic corresponding to a friction coefficient of a road surface and for producing a target yaw rate signal; deviation calculating means responsive to the target yaw rate and the actual yaw rate to calculate a yaw rate deviation; yaw rate gain determining means responsive to the vehicle speed signal for determining a yaw rate gain corresponding to a present vehicle speed by referring to a yaw rate gain map and for outputting a yaw rate gain signal; yaw moment calculating means responsive to the deviation signal and the yaw rate gain signal for deriving a yaw moment by using a predetermined moment equation of the vehicle and for producing a moment signal; and rear differential limiting torque calculating means responsive to the moment signal for calculating a limiting torque of the rear differential in response to the yaw moment by means of a predetermined equation to transmit a duty ratio pulse signal to the rear solenoid valve so as to perform an optimum control of the vehicle by controlling torque distribution to the left and right rear wheels and to improve stability and drivability when the vehicle is negotiating a tight a corner on a slippery road.

Further according to the present invention there is provided a torque distribution control method to distribute torque between the rear wheels of a vehicle said vehicle having an engine, a transmission connected to said engine via a clutch, a rear differential mechanically interposed between a left rear wheel and a right rear wheel for absorbing a left and right rear wheel speed difference, and a rear solenoid valve connected to said rear differential, comprising the steps of: sensing a steering angle of said front wheels; sensing a vehicle speed; sensing an actual yaw rate of said vehicle; calculating a target yaw rate from the sensed steering angle and the sensed vehicle speed; calculating a yaw rate deviation value of said calculated yaw rate from said actual yaw rate; determining a yaw rate gain from the vehicle speed, calculating a yaw moment by using a predetermined moment equation of said vehicle from said calculated yaw rate deviation and said yaw rate gain; and calculating a limiting torque of said rear differential by a predetermined equation to transmit an operating signal to said rear solenoid valve to perform an optimum control of said vehicle by distributing torque to said left and right rear wheels and to improve stability and drivability when said vehicle is negotiating a tight a corner on a slippery road.

In order that the invention may be more readily understood it will now be described, by way of example only, with reference to the accompanying drawings in which: Fig. 1 is a block diagram showing means constituting a torque distribution control system between front and rear wheels of a vehicle according to the present invention; Fig. 2 is a diagram showing the drive train of a four-wheel drive vehicle and a hydraulic control system thereof according to the present invention; Fig. 3 is a diagram showing a two-wheel vehicle model in the lateral motion; Fig. 4 is a block diagram showing means for estimating the friction coefficient of the road surface according to a second embodiment of the present invention; Fig. 5 is a diagram showing the relationship between the cornering power and the tire slip angle; Fig. 6 is a block diagram showing means constituting the torque distribution control between rear-left and rearright wheels of a vehicle according to the present invention; and Fig. 7 is a diagram showing a two-wheel vehicle model in the turning motion.

Referring now to Fig. 2, an outline of the construction of a power train for a four-wheel drive vehicle capable of controlling a torque distribution between front and rear wheels and between left and right wheels will be described hereinafter.

Numeral 1 denotes an engine, numeral 2 a clutch, numeral 3 a transmission, numeral 4 an output shaft of the transmission 3 and numeral 20 a centre differential. The output shaft 4 of the transmission 3 connects the transmission 3 and the centre differential 20. The centre differential 20 is connected through a front drive shaft 5 to a front differential 7 which drives a front-left wheel 9L and a front-right wheel 9R through a drive shaft 8.

Further, a rear drive shaft 6 and a propeller shaft 10 connect the centre differential 20 and a rear differential 11 which drives a rear-left wheel 13L and a rear-right wheel 13R through a drive shaft 12.

The rear differential 11 is composed of bevel gears and in this embodiment, a hydraulic type multiple disc rear clutch 28 as differential limiting means is provided between differential case ila and a side gear llb of the rear differential 11. In the case where the rear differential limiting torque of the rear clutch 28 is zero, the torque is equally distributed to the rear-left wheel 13L and the rearright wheel 13R. In the case where rear differential limiting torque is produced and becomes a value Td, the torque is shifted by the amount of the value Td from a high speed wheel to a low speed wheel and when a differential lock is engaged at the maximum value of the differential limiting torque Id, the torque is distributed to the two wheels 13L, 13R according to a product W y of a Load W applied to the rear-left wheel 13L and the rear-right wheel 13R respectively and of the friction coefficient p of road surface.

The center differential 20 is composed of compound type planetary gears and it comprises: a first sun gear 21 fixed to the output shaft 4 of the transmission 3, a second sun gear 22 fixed to the rear drive shaft 6, a plurality of pinion shafts 23 arranged around these sun gears 21, 22, a first pinion 23a fixed to the pinion shaft 23 and meshing with the first sun gear 21, and a second pinion 23b fixed to the pinion shaft 23 and meshing with the second sun gear 22.

Further, it comprises: a drive gear 25 rotatably mounted on the output shaft 4, a carrier 24 fixed to the drive gear 25 and rotatably connected to the pinion shaft 23 and a driven gear 26 fixed to the front drive shaft 5 and meshing with the drive gear 25. In the center differential thus constituted, the torque inputted to the first sun gear 21 is divided into the carrier 24 and the second sun gear 22 in a predetermined reference torque distribution ratio and the rotational dif ference generated between front and rear shafts when the vehicle turns a corner is absorbed by the planetary rotation of the pinion shaft 23. The reference torque distribution ratio can be determined at a desired value by the selection of the intermeshing pitch circles of the sun gears 21, 22 and the pinions 23a, 23b. Where et is the reference torque distribution ratio, TF is the front wheel torque and TR is the rear wheel torque, it is possible to establish the reference torque distribution ratio et as follows for example: TF : TR = 34 : 66 1 In this case, it is understood that the torque distribution ration has been set so as to apply a larger torque to rear wheels than front ones.

Further, a center clutch 27 of the hydraulic multiple friction disc type is disposed at the direct rear of the center differential 20. The center clutch 27 comprises a drum 27a coaxially fixed to the carrier 24 and a hub 27b coaxially fixed to the rear drive shaft 6. By controlling the center clutch 27 a differential limiting torque ic is produced so as to limit the differential operation of the center differential 20 and further it becomes possible to transfer the torque from rear to front wheels as well as from front to rear wheels.

In case of a front-engine, where WF denotes a front wheel weight, WR rear wheel weight and ew a static weight distribution ratio between the front wheel weight WF and the rear wheel weight WR, assuming ew as follows; WF : WR = 62 : 38 2 If the center clutch is fully engaged and friction coefficients of front and rear wheels against load surface are equal to each other, the torque is distributed between front and rear wheels as at the ratio expressed in the formula (2).

On the other hand, since the torque can be distributed also at the ratio expressed in the formula (1), the torque distribution ratio can be controlled widely between (1) and (2) according to the differential limiting torque Tc of the center clutch 27.

Next, the hydraulic control system for controlling the center clutch 27 and the rear clutch 28 will be de described.

Center clutch hydraulic control means comprise a hydraulic pump 30 for generating hydraulic pressure, a pressure regulator 31 for regulating hydraulic pressure, a hydraulic passage 33, a pilot valve 36 for further regulating hydraulic pressure, a hydraulic passage 38, an orifice 37, a duty solenoid valve 40 for producing a duty pressure Pd and a clutch control valve 34 for operating the center clutch 27 according to the duty pressure Pd. That is to say, the differential limiting torque Tc is variably controlled according to the magnitude of the duty pressure Pd.

On the other hand, rear clutch hydraulic control means 32' comprise a duty solenoid valve 40' for producing a duty pressure Pd and a clutch control valve 34' in addition to the hydraulic pump 31, the pilot valve 36 and others which are shared with clutch hydraulic control means. The differential limiting torque Td of the rear clutch 28 is variably controlled according to the duty pressure Pd in the same way as in aforementioned center clutch control means.

Next, the control of the torque distribution control system will be described. First, the torque distribution control between front and rear wheels will be described hereinafter.

When the tire characteristic is in a linear region, the cornering powers of front and rear wheels are constant but when the vehicle shows a marginal behavior such as spin due to a lost grip of tire while the vehicle turns a corner with acceleration on a road having a low friction coefficient, the tire lateral a t e ta @ force is lowered. The present invention is based on an idea that it is possible to estimate friction coefficients of road surface by treating the lowering of tire lateral force as the lowering of cornering power.

Furthermore, based on this idea it is possible to analyze the equation of vehicular motion extended to a non-linear region of the tire characteristic when the vehicle runs on a low friction coefficient road.

According to the theory of friction circle, it is known that the tire lateral force is affected by the driving force and that the stability of vehicle at the non-linear slip region can be judged from the stability factor of steering characteristic.

Thus, according to the present invention, first a friction coefficient of road surface is estimated by obtaining a cornering power of front and rear wheels in the nonlinear region based on miscellaneous parameters and a critical behavior of the vehicle is expressed numerically by use of the stability factor. Further, characteristics of vehicular motion in the linear region can be accurately grasped by analyzing equations of vehicular motion based on driving force, running conditions, friction coefficient of road surface and stability factor. Hence, it is possible to improve a stability of vehicle, such as preventing the vehicle from spinning, by controlling torque distribution between front and rear wheels so as to always obtain a constant staability factor.

Consequently, it is important to obtain the cornering power of front and rear wheels in the non-linear region based on miscellaneous parameters and to estimate a friction coefficient of road surface based on the cornering power. The cornering power can be obtained from the steering angle, the vehicle speed and the actual yaw rate. As a method of estimating the friction coefficient of road surface, there is a method in which, for example, the cornering power is estimated by comparing the yaw rate calculated from the equation of vehicular motion with the actual yaw rate on an on-line base.

Specifically, the cornering power is calculated by the method of parameter adjustment according to an adaptive control theory which will be described hereinafter.

First, an equation of vehicular lateral motion is produced using a vehicular motion model as illustrated in Fig. 3. The equation is expressed as: 2Cf + 2Cr = M-Gy 3 where Df, C r are the cornering force of the left and right wheels respectively, M is the vehicle weight and Gy is the lateral acceleration.

On the other hand, an equation of vehicular motion around the center of gravity is expressed as: 2Cf.Lf - 2Cr.Lr = Iz.# 4 where Lf, Lr are the distance from the center of gravity to the front and rear wheels respectivelY, Iz is the yaw moment of inertia of the vehicle and r is the yaw rate.

A lateral acceleration Gy is expressed as: Gy = Vy + V-r 5 where V is the vehicle speed and Vy is the lateral slip velocity.

Further, although the cornering forces Cf, Cr have a response like a first-order time Lag, if this time lag is neglected, the cornering forces Cf, Cr are expressed as: Cf = Kf af, Cr = Kr.αr 6 where Kf, Kr are the cornering power of the front and rear wheels respectively and a f, a r are the lateral slip angle of the front and rear wheels respectively.

On the other hand, when an idea of an equivalent cornering power is introduced in consideration of the effects of the rolling of the vehicle or the suspension system, the lateral slip angles a f, a r are expressed as: df Vy+Lfr af= n V 7 Vy-Lr r a r= 6 r- V where 8 f, 6 r. are the steering angle of front and rear wheels respectively and n is the steering gear ratio.

The abovementioned equations 3-7 are fundamental equations of motion.

Miscellaneous parameters are estimated by expressing these equations as variables of state and applying a parameter adjusting method to the adaptive control theory.

The cornering power is obtained from the parameters thus estimated. With respect to the parameters of an actual vehicle, there are vehicle weight, yaw moment of inertia and the Like. In developing the theory, these vehicular parameters are assumed to be constant and only the cornering power is assumed to be variable. The cornering power of the tire varies according to non-linerity of the lateral force against the slip angle, the effect of the friction coefficient of road surface, the effect of the weight transfer and the like. Where a is a parameter estimated by the change of the yaw rate r and b is a parameter estimated by the front wheel steering angle af, the cornering power of the front and rear wheels Kf, Kr are expressed as follows, for example: Kf = b.Iz.n / 2Lf 8 Kr = (a-Iz + Lf-Kf)/ Lr The cornering power of the front and rear wheels Kf, Kr in the non-linear region are estimated by substituting the vehicle speed V, the steering angle f and the yaw rate r into the aforementioned equations. Further, by comparing the cornering powers Kf, Kr thus estimated with those on a high friction coefficient road surface for each front and rear tire in the following manner for example, a friction coefficient of the road surface is caLcuLated and further based on the calculated friction coefficient p an estimated friction coefficient E in the non-linear region is determined with high accuracy.

f = Kf / KfO 9 r = Kr / KfO where f, r r are the friction coefficient of the front and rear wheels respectively, Kf, Kr; are the estimated cornering power of the front and rear wheels respectively, and KfO, KrO are the equivalent cornering power on the high friction coefficient road of the front and rear wheels respectively. Here, the equivalent cornering powers KfO, KrO are friction coefficients given by correcting the tire characteristic assumed to produce a cornering force in proportion to the slip angle within a region where a tire slip angle is very small by characteristics of suspension and others.

The above equations are understood as follows: When the vehicle operates on a high friction coefficient road with a full grip of tire, since both the front and rear wheels are employed in the linear region of tire characteristic, the estimated cornering powers Kf, Kr can be considered to be equal to the equivalent cornering powers KfO, KrO respectively and consequently the friction coefficients ps are estimated to be 1.0. Further, when the vehicle is drifting out, the slip angle of the front wheel becomes very large and accordingly, as shown in Fig. 5, it is estimated that the estimated cornering power Kf = cornering force of front wheel 1 slip angle of front wheel becomes extremely small. Similarly, when the vehicle is spinning, the estimated cornering power Kr = cornering force of rear wheel / slip angle of rear wheel becomes extremely small. To avoid this problem, the larger one of the estimated friction coefficients for front and rear wheels is let to be established as an estimated friction coefficient of the road surface "E".

Next, a case where the torque is distributed between the front and rear wheels will be described.

The equation of motion of a vehicle can be analyzed by extending them to the non-linear region using the vehicle speed V, the yaw rate T, the input torque Ti, the target stability factor At, the estimated friction coefficient E of road surface and others. The torque distribution ratio α between the front and rear wheels is calculated according to the following equations of motion of the vehicle.

Gx' =(Ti . Gt/Rt)/(W/g) Gy' =V . γ

where Gx' is the estimated longitudinal acceleration, Gy' is the estimated lateral acceleration, W isthe vehicle weight, e isthe height of a center of gravity, L isthe wheel base, Lf is the distance between the center of gravity and the front wheel, Lristhe distance between the center of gravity and the rear wheel, KfO, KrOarethe equivalent cornering power of the front and rear wheels in the linear region respectively, Kfc, Krc are theweight dependency of cornering power subjected to partial differential of Kf, Kr by the load of ground contact, Gt is the final gear ratio, Rt isthetire diameter, Ti isthe input torque, At is the target stability factor, AtO is the reference target stability factor (a predetermined constant, set at weak understeer), df is the steering angle of front wheel, GY is the yaw rate gain; nr is the deviation between act u a 1 yaw rate and target yaw r a t e, and V is the vehicle speed.

Based on the aforementioned equations, the control system shown in Fig. 1 will be described hereinafter.

Miscellaneous data, steering angle f detected by a steering angle sensor 42, vehicle speed V detected by a vehicle speed sensor 43, yaw rate r detected by a yaw rate sensor 44, engine apeed N detected by an engine speed sensor 45, accelerator pedal angle 9 detected by an accelerator pedal angle sensor 46 and gear position P detected by a gear position sensor 47 are inputted to the control unit 50.

In friction coefficients estimating means 51, the cornering powers Kf, Kr of the front and rear wheels are estimated based on the input data, steering angle f, vehi- cle speed V and actual yaw rate T according to the aforementioned adaptive control theory. The friction coefficients of road surface are calculated from the ratios of the estimated cornering powers Kf, Kr to the equivalent cornering powers, KfO, KrO on the road with a high friction coefficient (p = 1.0). Further, in order to avoid troubles such that estimated the cornering powers of the front wheels become extremely small when the front wheels are in a drift state, namely when the vehicle would not turn even with a turned steering wheel, or estimated cornering powers of the rear wheels become too small when the vehicle is spinning, a larger friction coefficient is chosen as an estimated friction coefficient E of road surface from among the friction coefficients of the front and rear wheels.

Further, in target yaw rate determining means 52, the target yaw rate T t is determined based on the input data, steering angle bf and vehicle speed V. The target yaw rate Qt and the actual yaw rate T are inputted to target steering characteristics determining means 53 where the target stability factor At of the steering characteristic is determined and corrected in accordance with the difference between both yaw rates r t and T- . Incidentally, the stability factor is determined so as to have a characteristic of slightly weak under-steer in average vehicles. Thus, when the vehicle spins or drifts out, the stability factor At is established numerically according to the change of the actual yaw rate f.

On the other hand, in input torque estimating means 5 4,. an engine outPut power Te is estimated from the input data, engine speed N and accelerator pedal opening angle and an input torque Ti of the center differential is calculated by multiplying the estimated engine power Te by a gear ratio g at the gear position P.

These data, vehicle speed V, actual yaw rate T input torque Ii, target stability factor At and estimated friction coefficient E are inputted to torque distribution ratio calculating means 55 where the torque distribution ratio a between the front and rear wheels is calculated by using the aforementioned equations. The torque distribution ratio a and the input torque Ti are inputted to differen tial limiting torque calculating means 56 where a center differential Limiting torque Tc is calculated according to the following equation: Tc = (a - Di) Ti where Di is a reference torque distribution ratio determined by the combination of planetary gears of the center differential 20 as described before. In this embodiment, the weight distribution between the front and rear wheels is biased on the rear wheels. If the weight distribution is biased on the front wheels, the above equation is amended as follows: Tc = (Di - a ) Ti In these equations, if the calculated center differential limiting torque Tc is negative, it is let Tc = O. The torque Tc thus calculated is converted into a torque signal in center differential limiting torque calculating means 56 and the torque signal is inputted to duty ratio converting means 57 wherein it is converted into a specified duty ratio D and this duty.ratio is outputted to the solenoid valve 40.

Next, an operation of this embodiment will be described.

First, the power of the engine 1 is inputted to the transmission through the clutch 2 and the converted power is inputted to the first sun gear 21 of the center differential 20. As described before, since the reference torque distribution ratio et has been set being biased on the rear wheels, the power is outputted to the carrier 24 and the second sun gear 22 with this torque distribution ratio.

Further, when the center clutch 27 is released (disengaged), the power is transmitted to the front and rear wheels with this distribution ratio et. As a result, the vehicle has a driving performance like that of a front engine-rear drive vehicle. Further, since the center differential 20 is free, the vehicle can be turned freely while the rotational difference between the front and rear wheels is absorbed therein.

Here, when the duty signal is outputted from the control unit 50 to the solenoid valve 40, the differential limiting torque Tc is produced by hydraulic control means 32. The torque Tc bypasses across the second sun gear and the carrier 24 and is transferred towards the front wheels. As a result of this, a greater amount of the torque is distributed to the front wheels than to the rear wheels, whereby the torque distribution biased on the front wheels is obtained.

While the vehicle is running, signals of the steering angle f, the vehicle speed V and the actual yaw rate r are inputted to the control unit 50 and the vehicle behavior is always watched. When the vehicle is running on the road with a high friction coefficient of road surface, the actual yaw rate r coincides approximately with the target yaw rate r t which is determined based upon the steer ing angle df and the vehicle speed V by the target yaw rate determining means 52. As a result, the stability factor At is set at a weak under-steer side and the steering characteris tic of the vehicle is always retained in a weak under-steer.

The estimated friction coefficients E calculated by the road friction estimating means 51 is transmitted to the torque distribution ratio calculating means 55 in which the torque distribution ratio a is caLcuLated based on the data, this calculated friction coefficients E, the vehicle speed V, the actual yaw rate T. , the stability factor At and the input torque Ti.

In a straight ahead operation of the vehicle, the torque distribution ratio rr is determined primarily based on the input torque Ti and the estimated longitudinal acceleration Gx'.

On the other hand, when the vehicle is turning corners, the torque distribution ratio a is determined primarily based on the vehicle speed V and the estimated lateral acceleration Gy' determined by the actual yaw rate Since the actual yaw rate r is subjected to the feedback control, the control system according to the present invention is not affected by disturbances or control errors.

Next, when the vehicle is turning on the road with a low friction coefficient of road surface, if the torque is distributed more to the rear wheels than to the front wheels, first the lateral force of tire is decreased at the rear wheel side due to an excessive traction of the rear wheels and resultantly the rear wheels slip in the lateral direction. Eventually, when the tire grip exceeds a limit and the vehicle starts spinning, in the road friction estimating means 51 the cornering powers Kf, Kr of the front and rear wheels are estimated based on the steering angle af, the vehicle speed V and the actual yaw rate r in response to the vehicle behavior. Further, for each of the front and rear wheels the friction coefficient of road surface is calculated by comparing it with the one of the road with high friction coefficient and the highest one is selected from among friction coefficients thus calculated. This highest friction coefficient is the estimated friction coefficient E.

Further, in target steering characteristic determining means 53 the target stability factor At is determined according-to the formula 15 described before based on the yaw rate gain Gr determined by the yaw rate gain determining means 58 and on the deviation Ar of the actual yaw rate r from the target yaw rate rt determined by the target yaw rate determining means 52. For example, when the actual yaw rate becomes larger than the target yaw rate as a result of vehicular spin, the target stability factor At becomes larger than the reference target stability factor AtO, namely the target stability factor At is amended in the direction of strengthening under-steer. Then, in the torque distribution ratio calculating means 55 the torque distribution ratio a is calculated and controlled being biased on the front wheels and as a result of this the lateral force of the rear wheels increases, whereby the vehicle is prevented from spinning.

Thus, the. feedback control is performed through the target stability factor At such that the actual yaw rate coincides with the target yaw rate and accordingly the vehicle behavior is always kept at a favorable weak under-steer.

Next, the torque distribution control between left and right wheels will be described hereinafter. In this embodiment, as an example of the control between left and right wheels the torque distribution control between the rear-left and rear-right wheels will be explained.

The torque distribution control system according to this embodiment is based on the following principle: When a rear differential limiting torque Td increases during a high speed turn with an accelerator pedal released, the braking force of the outer rear wheel becomes larger than that of the inner rear wheel and, as a result, the difference of these braking forces generates a moment M for intending to drive the vehicle straight ahead. It is known that this moment M is effective for preventing a tackin of the vehicle. On the other hand, a magnitude of tack-in can be judged from the deviation of the actual yaw rate T from the target yaw rate r t which a vehicle driver determines according to the vehicle speed V and the steering angle df during turning. Where the numerated magnitude of tuck-in is identified as a variation of the stability factor, the tack-in can be prevented by generating a yaw moment M so as to cancel this variation of the stability factor. That is to say, in order to prevent the tack-in, the rear differential limiting torque Td should be determined so as to generate the yaw moment M.

First, the target yaw rate rt will be determined as follows.

In a two-wheels vehicle model as shown in Fig. 7, equations of motion are expressed as: mv(,8+t)'=Cf+Cr Ir=Lf- Of-Lr Cr 16 where 7 is the actual yaw rate, d is the slip angle of the vehicle, V is the vehicle speed (V is constant), m is the vehicle mass, I is the yaw moment of inertia, C f, Cr are cornering force of the front and rear wheels respectively, Lf, Lr are the distance between the center of gravity and an axis of the front and rear wheels respectively.

The relationship between the cornering force and the slip angle of tire in the linear region is expressed as Cf = 2 Kf a f, Cr = 2 Kr a' r, where Kf, Kr are an equivalent cornering power of the front and rear wheels respectively and af, a r are a slip angle of tire of the front and rear wheels respectively.

Introducing the above relationship into the equa tions 16, the equations of motion are further expressed as: m v ss + 2 (Kf + Kr) ss + {mv + 2 (LfKf - LrKr)/V} γ = 2Kf 6 f+ 2Kr 6 r 2 (LfKf - LrKr) ss + I γ + {2 (Lf Kf + Lr Kr)/V} γ = 2LfKf#f - 2LrKr#r 17 Based on the above fundamental equations of vehicular motion, the target yaw rate r t is obtained.

Next, it will be explained how the yaw moment M and the rear differential limiting torque Td are calculated.

When the rear differential limiting torque M is introduced into equations of motion of the two-wheels vehicle model shown in Fig. 7, they are expressed as: mv (ss + γ) = Cf + Cr 18 17 =LfCf-LrCr-M where T is the yaw rate (variable), @ is the slip angle of the vehicle (variable), m is the vehicle mass, V is the vehicle speed, Cf, Cr are the yaw moment of inertia, Lf, Lr are the distance spectively, ; is the yaw mojment of inertia, Lf, Lr are the distance between the center of gravity and an axis of the front and rear wheels respectively, M is the moment b y the rear di ffere n- tial limiting torque.

The cornering force of the front and rear wheels is expressed respectively as follows: Cf=2Kfa f 19 Cr=2Kra r where Cf, Cr are the cornering force of the front and rear wheels respectively, Kf, P are the cornering power of the front and rear wheels respectively, a f, a r are the slip angle of tire of the front and rear wheels respectively.

Where the steering angles f and #r are substi- tuted inio the tire slip angles a f and a r respectively, substitution of the equation 19 into the equations 18 gives the following m v ss + 2(Kf + Kr)ss + {m v + 2 (LfKf - LrKr)/V} γ =2Kf6f+2Kr6 r 2(LfKf - LrKr) ss + I γ + {2 (LfKf + Lr 2Kr)/V) 7 +M = 2LfKf#f - 2LrKr#r Next, when the vehicle turns around a fixed circle the characteristic thereof will be explained. In this case, the slip angle ss of the vehicle and the yaw rate T are both constant and therefore deviations thereof are let to be zero.

The equations 20 are expressed as follows: 2 (Kf + Kr) ss + { m v + 2 (LfKf - LrKr)/v} γ = 2Kf#f 2 (LfKf - LrKr) ss + { 2 (LfKf + LrKr)/v} γ + M = 2LfKf#f 21 where the steering angle # f o f t h e r e a r w h e e l should be zero.

The above equation is transformed into the following equation: 2 (LfKf - LrKr) ss + {2 (LfKf + LrKr) /v + M/ γ } / γ = 2LfKf 6 f 22 - The solution r is given by the equations 21 and 22 as follows: (Formula 23)

where L is the wheelbase (Lf + Lr), T in the right side; a yaw rate previously obtained.

In order for the equation 23 to have a physical meaning, it is necessary to meet the following condition:

Now introducing a stability factor A' extendedly applied to the vehicle with a rear differential limiting control, the following equation is given: Kf+Kr M A' = A + # 25 2LKfKr Vγ where A; a stability factor of a case where the rear differential limiting torque control is free.

Accordingly, when the yaw rate r is increased (Ar > o) by tack-in, the deviation AA of the stability factor is expressed as:

In the above equation, GT denotes a yaw rate gain of the steering angle # f of the front wheel and the yaw rate gain is expressed as follows:

Consequently, the moment M necessary for canceling the tack-in is expressed as follows: 2LKfKrV γ M = - # A Kf + Kr 28 Further, the rear differential limiting torque Td is given as: Td = (M / d) R 29 where R is the tire diameter, d is the tread.

Next, referring to Fig. 6, the function of the torque distribution control system will be described.

Signals of the yaw rate T detected by the yaw rate sensor 44, the steering angle a f detected by the steering angle sensor 42 and the vehicle speed V detected by the vehicle speed sensor 43 are inputted to a control unit 70. In the control unit 70, there are provided yaw rate gain determining means 71 in which a yaw rate gain Gr of the predeter mined steering angle # # f of the front wheel is determined from the aforementioned equations or by reading a map. The vehicle speed V and the steering angle a f are inputted to target yaw rate calculating means 72 where a target yaw rate r t corresponding to the running condition on the high friction coefficient road is calculated based upon the aforementioned equations of motion. The calculated target yaw rate γ t and the actual yaw rate T are inputted to deviation calculating means 62 where the deviation #γ (Ar = r - r γt: Ar > O) is calculated. Thus, a tuck-in is detected by an increase of the actual yaw rate T and further a magnitude of the tuck-in is obtained from the deviation Ar.

The yaw rate gain GT and the yaw rate deviation hT corresponding to the magnitude of tack-in are inputted to yaw moment calculating means 74. In the yaw moment calculating means 74, first a deviation of the stability factor is obtained as AA by using the yaw rate gain GT and the yaw rate deviation Ar. Since the stability factor is predetermined on the weak under-steer side, when a yaw rate deviation Ar is generated by tack-in, the deviation AA of the stability factor becomes a negative value (namely, on the oversteer side) according to the yaw rate deviation Ar. Based on the calculated deviation AA, finally a yaw moment M necessary for canceling the deviation #A is calculated.

The yaw moment M is inputted to rear differential limiting torque calculating means 75 where a rear differential limiting torque Td is calculated. Further, this torque signal Td is converted into a duty ratio D in duty ratio converting means 76 and then the duty signal D is outputted to the solenoid valve 40'.

The torque distributed by the center differential 20 and the center clutch 27 is transmitted to the rear differential 11. When the rear clutch 28 is released, the rear differential 11 distributes the driving force equally to the rear-left wheel 13L and the rear-right wheel 13R. Further, in this case, when the accelerator pedal is released. the braking force is also distributed equally. When the rear clutch 28 is engaged by the hydraulic control means 32', a differential limi-ting torque Td is generated in the rear clutch 28 and the torque distribution between the rear-left and rearright wheels, 13L and 13R is changed by the differential limiting operation. That is to say. in a case where the driving force is applied, the torque is transferred from the high speed wheel to the low speed wheel (grip wheel) according to the rear differential limiting torque Td. On the other hand, in a case where the rotational speed of the outer wheel is larger than that of the inner wheel while the vehicle turns with the accelerator pedal released, the braking force is distributed more to the outer wheel than to the inner wheel according to the rear differential limiting torque Td.

During operation of the vehicle. signals of the steering angle a f, the vehicle speed V and the yaw rate T are inputted to the control unit 70 and the vehicle behavior is always watched. When the vehicle behavior does not change while the vehicle runs straight ahead or turns corners, the target yaw rate Tt which has been calculated from the steering angle a f and the vehicle speed V in the target yaw rate calculating means 72 coincides with the actual yaw rate r and therefore the stability factor does not change. Accord ingly, the rear differential limiting torque Td remains to be zero.

On the other hand, when the vehicle enters into a tack-in, namely. when the vehicle turns abruptly inside, while the vehicle turns at high speed with the accelerator pedal released. the actual yaw rate 7 increases. Then, the deviation calculating means 73 calculates the deviation Ar of the actual yaw rate r from the target yaw rate r t and the magnitude of tack-in is detected. Further, in the yaw moment calculating means 74 this deviation AT is converted into the deviation AA of the stability factor and the yaw moment M for canceling this deviation AA is calculated therein.

Then, in the rear differential limiting torque calculating means 75, the rear differential limiting torque Td corresponding to the calculated yaw moment M is calculated and this torque Td is applied to the rear clutch 28. Thus, when the vehicle turns at high speed with the accelerator pedal released. according to this torque Td the braking torque is distributed more to the rear-outer wheel than to the the rear-inner wheel and as a result the moment M cancel ing the tack-in is generated so as to prevent the tack-in phenomenon. Further, since in the control system according to this embodiment a feedback control is perfomed such that the actual yaw rate T coincides with the target yaw rate r t, the vehicle never goes to a strong under-steer side adversely and only tack-in phenomenon can be securely prevented. Further, since the control system is constituted such that the yaw rate deviation Ar is converted into the deviation AA of the stability factor, the steering characteristic retains a weak under-steer as initially designed.

In this embodiment of the torque distribution control system, an example of the torque distribution control system between the rear-left and rear-right wheel has been described. However. fundamental features of the control system can be applied to a torque distribution control system between front-left and front-right wheel.

Referring to Fig. 4 a second embodiment according to the present invention will be described. The second embodiment according to the present invention is another example of the friction coefficients estimating means 51.

The principle of the second embodiment is based on estimating the cornering power of the front and rear wheels by extending the cornering power to a non-linear region according to the adaptive control theory whose variables comprise a deviation of lateral acceleration and a deviation of actual yaw rate. That is to say, in the adaptive observa tion system constituting the steering angle. the vehicle speed and the estimated cornering power, the cornering power is estimated by extending the yaw rate and the lateral acceleration to the non-linear region on the base of the vehicular motion model in the linear region.

The vehicle speed V detected by the vehicle speed sensor 43, the steering angle γf detected by the steering angle sensor 42 the yaw rate r detected by the yaw rate sensor 44 and the lateral acceleration Gy detected by the lateral acceleration sensor 48 are inputted to the friction coefficients estimating means 51. The friction coefficients estimating means 51 has yaw rate and Lateral acceleration calculating means (adaptive observation system) 61 to which the steering angle a f, the vehicle speed V and the cornering power Kf, Kr of the front and rear wheels are inputted. In the yaw rate and lateral acceleration calculating means 61 a yaw rate r n and a lateral acceleration Gyn are calculated by using these parameters according to the vehicular motion model in the linear region. The calculated yaw rate r n, the calculated lateral acceleration Gyn, the detected yaw rate r and the detected lateral acceleration Gy are inputted to deviation calculating means 62 in which a deviation AT of the calculated yaw rate Tn from the detected yaw rate r and a AG of the calculated lateral acceleration Gyn from the detected lateral acceleration Gy are calculated.

These deviations Ar and AG are inputted to tire characteristic control means (adaptive controller) 63 where the cornering powers Kf, Kr of the front and rear wheels in a marginal behavior are estimated. Here, in a case where the actual lateral acceleration Gy is decreased and AG is posi tive, since it is judged that the vehicle is drifting out or spinning in the marginal area, both the cornering powers Kf and Kr should be reduced. On the other hand, in a case where AG is negative, since it is judged that the vehicle is in tuck-in, both Kf and Kr should be increased. In a case where the actual yaw rate r is reduced and Ar is positive, judging that the vehicle is drifting out, the cornering power Kf of the front wheels should be reduced and Kr of the rear wheels should be increased. In a case where the actual yaw rate r is increased and t r is negative, judging that the vehicle is spinning, Kf of the front wheels should be increased and Kr of the rear wheels should be reduced. How the cornering powers Kf, Kr are corrected according to the state of both deviations Ar, AG is summerized in the following Table 1: Table 1 Kf Kr Reduce Reduce If AG > O Increase Increase If AG < O Reduce Increase If L r > 0 Increase Reduce If A r < O As shown in Fig. 5, the cornering powers Kf, Kr corresponding to drift-out or spin of the vehicle in the marginal region are determined accurately every moment by reducing or increasing the cornering powers previously obtained by a predetermined increment according to the Table 1.

Thus estimated cornering powers Kf, Kr of the front and rear wheels are inputted to friction coefficients determining means 64 and the friction coefficient of the front and rear wheels are estimated respectively by comparing the estimated cornering powers with those of the high friction coefficient road in the same manner as in the first embodiment. The estimated friction coefficient E is a larger one among those friction coefficients estimated above.

In this second embodiment, the yaw rate and lateral acceleration calculating means 61 of the friction coefficients estimating means 51 calculate the yaw rate rn and the lateral acceleration Gyn based on the steering angle df the vehicle speed V, the estimated cornering powers Kf, Kr at the previous moment and the deviation calculating means 62 calculate the deviation A r of the actual yaw rate r from the calculated yaw rate Tn and the deviation AG of the actual lateral acceleration G y from the calculated lateral accele-ration Gyn. Further, the tire characteristic control means 63 estimate the cornering powers Kf, Kr at the present moment based on the correction of the cornering power according to the state of both deviations Ar and AG. When the vehicle drifts out or spins on the low friction coefficient road, the vehicle behavior is detected as the deviation Ar of yaw rate and the deviation AG of lateral acceleration and the cornering powers Kf, Kr of the front and rear wheels in a sideslip state can be estimated with further high accuracy.

In summary, the torque distribution control system according to the present invention provides a safe and comfortable driving under any conditions of road or in a marginal state of behavior by properly distributing the driving torque among wheels.

The torque distribution control mechanism between the front and rear wheels comprises friction coefficients estimating means for estimating a friction coefficient of road surface from cornering powers of the front and rear wheels, target yaw rate determining means for determining a target yaw rate according to the running condition of the vehicle on the high friction coefficient road based on a steering angle and a vehicle speed, target steering characteristic determining means for determining a target stability factor according to the difference between the target yaw rate and the actual yaw rate, input torque estimating means for estimating an input torque of the center differential, torque distribution ratio calculating means for calculating a torque distribution ratio between the front and rear wheels from equations of motion of the vehicle extended to a nonlinear region, based on the vehicle speed. the actual yaw rate, the input torque, the estimated friction coefficient of road surface and the target stability factor. and center differential limiting torque calculating means for calculating a center differential limiting torque based on the torque distribution ratio between the front and rear wheels and the input torque.

In thus constituted torque distributing control mechanism between the front and rear wheels, when the vehicle operates on a road with high friction coefficients. the torque is properly distributed between the front and rear wheels according to the running conditions of the vehicle such as a straight running and a cornering operation, whereby an excellent maneuverability is given to the vehicle.

On the other hand, when the vehicle operates on a road with low friction coefficients, since a friction coefficient can be estimated with high accuracy and the torque distribution is properly distributed between the front and rear wheels, the vehicle can be prevented from spinning or drifting out.

The torque distributing mechanism between the rear-left and rear-right wheel according to the present invention comprises target yaw rate calculating means for calculating a target yaw rate from equations of vehicular motion based on the steering angle and the vehicle speed, deviation calculating means for calculating a deviation of the actual yaw rate from the target yaw rate according to the magnitude of tack-in, yaw rate gain determining means for determining a yaw rate gain based on a predetermined map parameterizing the vehicle speed, yaw moment calculating means for calculating a yaw moment necessary for canceling the deviation of the stability factor which has been calculated from the above deviation of the actual yaw rate and the above yaw rate gain, and rear differential limiting torque calculating means for calculating a rear differential limiting torque according to the above yaw moment.

In thus constituted torque distributing mechanism between the rear-left and the rear-right wheel, since the magnitude of tack-in can be accurately detected. a rear differential limiting torque necessary for canceling the tack-in can be calculated. The calculated rear differential limiting torque generates a yaw moment, whereby a tack-in phenomenon of the vehicle can be prevented.

While the presently preferred embodiments of the present invention have been shown and described, it is to be understood that these disclosures are for the purpose of illustration and that various changes and modifications may be made without departing from the scope of the invention as set forth in the appended claims.

Claims

CLAIMS 1. A torque distribution control system for distributing an torque to the wheels of the vehicle, the vehicle having an engine, a transmission connected to the engine via a clutch, a rear differential interposed between a left rear wheel and a right rear wheel to absorb a left and right rear wheel speed difference, a rear solenoid valve hydraulically connected to the rear differential, the system comprising: a steering sensor mounted on the vehicle for sensing a steering angle of the front wheels and for generating a steering angle signal, a vehicle speed sensor mounted on the vehicle for detecting vehicle speed and for producing a vehicle speed signal, a yaw rate sensor mounted on the vehicle for detecting a yaw rate of the vehicle and for generating an actual yaw rate signal: a target yaw rate determining means responsive to the steering angle signal and the vehicle speed signal for setting a target yaw rate based upon a standard turning characteristic corresponding to a friction coefficient of a road surface and for producing a target yaw rate signal; deviation calculating means responsive to the target yaw rate and the actual yaw rate to calculate a yaw rate deviation; yaw rate gain determining means responsive to the vehicle speed signal for determining a yaw rate gain corresponding to a present vehicle speed by referring to a yaw rate gain map and for outputting a yaw rate gain signal; yaw moment calculating means responsive to the deviation signal and the yaw rate gain signal for deriving a yaw moment by using a predetermined moment equation of the vehicle and for producing a moment signal; and rear differential limiting torque calculating means responsive to the moment signal for calculating a limiting torque of the rear differential in response to the yaw moment by means of a predetermined equation to transmit a duty ratio pulse signal to the rear solenoid valve so as to perform an optimum control of the vehicle by controlling torque distribution to the left and right rear wheels and to improve stability and drivability when the vehicle is negotiating a tight a corner on a slippery road.
2. A torque distribution control system as claimed in claim 1 and as herein described with reference to the accompanying drawings.
3. A torque distribution control method to distribute torque between the rear wheels of a vehicle said vehicle having an engine, a transmission connected to said engine via a clutch, a rear differential mechanically interposed between a left rear wheel and a right rear wheel for absorbing a left and right rear wheel speed difference, and a rear solenoid valve connected to said rear differential, comprising the steps of: sensing a steering angle of said front wheels; sensing a vehicle speed; sensing an actual yaw rate of said vehicle; calculating a target yaw rate from the sensed steering angle and the sensed vehicle speed; calculating a yaw rate deviation value of said calculated yaw rate from said actual yaw rate; determining a yaw rate gain from the vehicle speed, calculating a yaw moment by using a predetermined moment equation of said vehicle from said calculated yaw rate deviation and said yaw rate gain; and calculating a limiting torque of said rear differential by a predetermined equation to transmit an operating signal to said rear solenoid valve to perform an optimum control of said vehicle by distributing torque to said left and right rear wheels and to improve stability and drivability when said vehicle is negotiating a tight a corner on a slippery road.

A A method of controlling torque distribution as claimed in claim 3 and as herein described.