DE19549715B4 - Torque distribution control system for four wheel drive motor vehicle - controls torque distribution ratio to wheels in accordance with estimate of road surface friction coefficient, calculated based on detected steering angle, vehicle speed and yaw rate - Google Patents

Abstract

The torque distribution control system for distributing an input torque to the wheels of a vehicle comprises a road friction estimating device for estimating a friction coefficient of the road surface, a torque distribution ratio calculating device for calculating a torque distribution ratio by using the friction coefficient, and a torque distributing mechanism for distributing the input torque among the wheels based on the torque distribution ratio. The torque distribution control system further includes yaw moment calculating device for calculating a yaw moment, and a torque distribution ratio calculating device for calculating a torque distribution ratio according to the yaw moment. A torque distribution mechanism for distributing the input torque to the wheels based on the torque distribution ratio.

Description

The invention relates to a method to determine the road friction between the road and the tire for a Control system of a motor vehicle.

It is well known that a vehicle one for his respective drive system - for example a front engine / rear-wheel drive system (FR system), a front engine / front-wheel drive system (FF system) - specific Shows driving behavior. Moreover it is known that a Vehicle with permanent all-wheel drive, which is an intermediate differential has, in terms of driving behavior in the border area compared to ordinary FR or FF vehicles can be improved if, for example suddenly is braked or when the vehicle is cornering. Such Four-wheel drive vehicles are becoming increasingly popular because of their steering behavior between over and understeer and therefore it is said that four wheel drive vehicles and intermediate differential are easy to drive.

An example of a control system for distributing driving force between the front and rear wheels of a four-wheel drive vehicle and an intermediate differential is the unexamined JP patent application JP 63-13824 A , in which a lateral acceleration is detected during cornering and a differential limiting torque is generated in accordance with the extent of the lateral acceleration in a hydraulic multi-plate clutch, so that the drive force distribution between the front and rear wheels is controlled so that no spinning or pushing outward during cornering is effected.

Further examples of this technique can be found in the unexamined JP patent applications JP 61-229616 A and JP 3-74221 A , The former document shows a technique in which the driving force distribution between front and rear wheels or between the left and right wheels is variably controlled by spinning or pushing out based on a difference between a target yaw angle, the is calculated from the steering angle and the vehicle speed, and an actual yaw angle is detected, and the latter document shows a technique in which the driving force distribution between front and rear wheels or between the left and right wheels is variably controlled by changing the steering characteristic is calculated over time from a steering angle, a vehicle speed and the actual yaw angle.

In these known techniques, for example according to the JP 63-13824 A However, since the cornering state is only detected by the lateral acceleration, the controllable range is limited to a so-called linear tire grip range, in which the lateral force changes in proportion to a slip angle of the tire. Thus, when the vehicle runs on a road with a small coefficient of static friction or friction and reaches such a limit range that the tire holding force reaches a limit and the vehicle's wheels start to spin, the lateral force changes in a non-linear manner and the actual lateral acceleration changes arbitrarily in accordance with the vehicle behavior in a spinning state, so that a cornering state of the vehicle cannot be assessed precisely. Because also in the specified documents JP 61-229616 A and JP 3-74221 A If the cornering state of the vehicle is judged on the basis of a signal from a yaw angle sensor, it can be expected that the vehicle behavior will be more precise than in the document JP 63-13824 A can be judged, but this state of the art is still insufficient in the control in the limit state.

Furthermore, from the DE 39 12 014 A1 A method for determining the coefficient of friction between the road surface and the tires of a vehicle is known, the coefficient of friction being determined from the steering angle, the vehicle speed, the yaw rate and the lateral acceleration.

The invention is intended to address those discussed above Disadvantages and shortcomings remove the state of the art.

The object of the invention is to provide a further developed procedure for the determination of the road friction between the road and the tires for a control system in one Motor vehicle.

This task is solved according to the teaching of claim 1.

The driving force distribution control system, which can be operated with the method according to the invention, has: a speed sensor for measuring an engine speed; an accelerator opening angle sensor for measuring an accelerator opening angle; a gear position sensor for detecting a gear position; a steering angle sensor for measuring a steering angle; a vehicle speed sensor for measuring a vehicle speed; a yaw rate sensor for measuring a yaw rate of the vehicle; input torque estimating means for estimating an input torque based on the engine speed, the accelerator opening angle and the transmission position; a target yaw angle determination direction for determining a target yaw angle based on the steering angle and the vehicle speed; yaw rate increase determining means for determining a yaw rate increase based on the vehicle speed; target steering characteristic determining means for determining a target stability factor based on the target yaw rate, yaw rate, vehicle speed, and yaw rate increase; road surface friction estimation means for estimating the friction coefficient of a road surface based on the steering angle, the vehicle speed and the yaw angle; and a driving force distribution ratio calculator for calculating a driving force distribution ratio based on the input torque, the target stability factor, the vehicle speed, the yaw angle, and the coefficient of friction.

The invention is illustrated below the description of exemplary embodiments and explained in more detail with reference to the accompanying drawings. The Drawings show in:

1 Fig. 3 is a block diagram showing devices constituting a driving force distribution control system for distributing driving force to front and rear wheels of a vehicle;

2 4 is a diagram showing the power transmission of a four-wheel drive vehicle and a hydraulic control system therefor according to the invention;

3 a diagram of a two-wheel vehicle model in the lateral movement;

4 a diagram of the relationship between the lateral force coefficient and the slip angle of the tire;

5 a block diagram showing the means for controlling the distribution of driving force to the left and right rear wheels of a vehicle; and

6 a diagram of a two-wheel vehicle model in the cornering movement.

With reference to 2 The principle of the structure of the power transmission for a four-wheel drive vehicle is described below, in which the control of the driving force distribution on the front and rear wheels and on the left and right wheels is possible.

A motor is provided 1 , a clutch 2 , a gear 3 , an output shaft 4 of the transmission 3 and an intermediate differential 20 , The output shaft 4 of the transmission 3 connects the gearbox 3 with the intermediate differential 20 , The intermediate differential 20 is via a front drive shaft 5 with a front axle differential 7 connected, which is a left front wheel 9L and a right front wheel 9R through a drive axle 5 drives. A rear drive shaft 6 and a propeller shaft or propeller shaft 10 connect the intermediate differential 20 and the rear axle differential 11 with each other that's a left rear wheel 13L and a right rear wheel 13R via a drive axle 12 drives.

The rear axle differential 11 consists of bevel gears and in this embodiment is a hydraulic rear axle multi-plate clutch 28 as a differential limiting unit between a differential case 11a and an axle shaft bevel gear 11b of the rear axle differential 11 intended. If a rear axle differential limiting torque of the rear axle clutch 28 Is zero, the torque is even on the left rear wheel 13L and the right rear wheel 13R is distributed, and when the rear differential differential torque is generated and becomes Td, the torque is redistributed by the value Td from a high-speed wheel to a low-speed wheel, and when a differential lock becomes effective at the maximum value of the differential limitation torque Td, the torque becomes the two wheels 13L . 13R according to a product W – μ from a load W applied to the left rear wheel 13L or the right rear wheel 13R is applied, and distributed a coefficient of friction μ of the road.

The intermediate differential 20 is built from a compound planetary gear and has the following: a first sun gear 21 that on the output shaft 4 of the transmission 3 is attached, a second sun gear 22 that on the rear drive shaft 6 attached, a variety of differential wheel axles 23 that around these sun gears 21 . 22 are arranged around a first differential gear 23a that on the differential wheel axle 23 is attached and with the first sun gear 21 combs, and a second differential gear 23b that on the differential wheel axle 23 is attached and with the second sun gear 22 combs.

The planetary gear also has: a drive wheel 25 that on the output shaft 4 is rotatably attached to a planet carrier 24 on the drive wheel 25 attached and with the differential wheel axle 23 is rotatably connected, and a driven wheel 26 that on the front drive shaft 5 is attached and with the drive wheel 25 combs. In the intermediate differential thus constructed, the input torque becomes the first sun gear 21 in a predetermined reference torque distribution ratio on the planet carrier 24 and the second sun gear 22 split, and the rotation difference generated between the front and rear axle shafts when the vehicle is cornering is determined by the planetary rotation of the differential gear axis 23 absorbed. The reference torque distribution ratio can be determined with a desired value by selecting the partial circles of the sun gears which are in engagement with one another 21 . 22 and the differential gears 23a . 23b , If it is the reference torque distribution ratio, TF is the front wheel torque and TR is the rear wheel torque, then it is possible to determine the reference torque distribution ratio et as follows, for example:

(Formula 1):

• TF: TR = 34: 66

In this case, it goes without saying that this Torque distribution ratio is set so that on the rear wheels a greater torque than on the front wheels is applied.

A central clutch designed as a multi-plate hydraulic friction clutch 27 is the intermediate differential 20 immediately downstream. The central coupling 27 includes a drum 27a that are coaxial on the planet carrier 24 is attached, and a hub 27b that are coaxial on the rear drive shaft 6 is attached. By controlling the central clutch 27 a differential limiting torque Tc is generated to compensate for the intermediate differential 20 limit, and it also becomes possible to transfer the torque from rear wheels to front wheels and from front wheels to rear wheels.

In the case of a front engine, where WF is a front wheel weight, WR is a rear wheel weight and ew is a static weight distribution ratio between the front wheel weight WF and the rear wheel weight WR, ew should be assumed as follows:

(Formula 2):

• WF: WR = 62: 38.

When the central clutch is fully engaged and the friction values of the front and rear wheels on the road are equal to each other, the torque is distributed between the front and rear wheels with the ratio expressed in the formula (2). However, since the torque can also be distributed with the ratio expressed in the formula (1), the torque distribution ratio can be in a wide range between (1) and (2) in accordance with the differential limiting torque Tc of the central clutch 27 to be controlled.

Next are the hydraulic control system for controlling the central clutch 27 and the rear axle coupling 28 described.

The hydraulic control device for the central clutch comprises a hydraulic pump 30 to generate hydraulic pressure, a pressure regulator 31 to regulate the hydraulic pressure, a hydraulic line 33 , an auxiliary control valve 36 for further regulation of the hydraulic pressure, a hydraulic line 38 , a choke 37 , an operating solenoid valve 40 for generating an operating pressure Pd and a clutch control valve 34 to operate the central clutch 27 according to the operating pressure Pd. That is, the differential limiting torque Tc is variably controlled in accordance with the value of the operating pressure Pd.

On the other hand, the hydraulic control device includes 32 ' an operating solenoid valve for the rear axle coupling 40 ' for generating an operating pressure Pd and a clutch control valve 34 ' in addition to the hydraulic pump 31 , the auxiliary control valve 36 and other elements that are provided together with the hydraulic clutch control device. The differential limiting torque Td of the rear axle clutch 28 is variably controlled in accordance with the operating pressure Pd in the same manner as in the control device for the central clutch.

Next is the control of the Driving force distribution control system described, first controlling the driving force distribution on the front and rear wheels explained becomes.

If the tire characteristics located in a linear range are the lateral force coefficients the front and rear wheels constant, but if the vehicle has a limit behavior such as "spinning" due to a lost one Tire grip or grip shows while the vehicle is accelerating a curve on a road drives through with a low coefficient of friction, becomes the cornering leader of the tire is reduced. The tax system is based on the idea that it possible is to estimate road friction values by reducing the cornering leader of the tire is treated as a reduction in the lateral force coefficient becomes. On the basis of this idea it is also possible to work on a non-linear Tire characteristic extended vehicle motion equation to analyze when the vehicle is on a small lane Friction coefficient drives.

According to the theory of the friction circle it is known that the Cornering force of the tire is influenced by the driving force and the stability of a tire Vehicle in the non-linear slip area due to the stability factor can be judged by steering characteristics.

Therefore, first a coefficient of friction the roadway estimated by obtaining a lateral force coefficient of front and rear wheels in a nonlinear range based on different ones Parameters, and a critical behavior of the vehicle is numerically determined by Application of the stability factor expressed. Moreover can Characteristics of the vehicle movement can be precisely recorded in the linear range by analyzing vehicle motion equations based on that the driving force, driving conditions, the coefficient of friction of the road and the stability factor. It is therefore possible the stability to improve the vehicle, for example a spinning of the Prevent vehicle by distributing the driving force to the vehicle Front and rear wheels is controlled so that always a constant stability factor is obtained.

It is therefore important that Lateral force coefficient of the front and rear wheels in the non-linear area based on different parameters and get one To estimate the coefficient of friction of the road surface on the basis of the lateral force coefficient. The Lateral force coefficient can be determined from a steering angle, a vehicle speed and an actual yaw angle can be obtained. In a method for Estimate A coefficient of friction of the road becomes, for example, the lateral force coefficient estimated by comparing that calculated from the vehicle motion equation Yaw angle with the actual yaw angle on an online basis. The lateral force coefficient is used according to the method of parameter setting according to a theory the adaptive controller, as will be described.

First, an equation of vehicle lateral motion is formed using a vehicle motion model as shown in FIG 3 is shown. The equation is written as follows:

(Formula 3):

• Cf + 2Cr = M · Gy, with Cf, Cr = cornering force of the left or right wheel; M = vehicle weight; and Gy = lateral acceleration.

On the other hand, it becomes an equation the vehicle movement around the center of gravity as follows:

(Formula 4):

• 2Cf · Lf - 2Cr · Lr = Iz · γ., with Lf, Lr = distance from the center of gravity to the front or rear wheel; Iz = yaw moment of inertia of the vehicle; and γ = yaw angle.

The lateral acceleration Gy becomes like written as follows:

(Formula 5):

• Gy = V .y + Vγ, with V = vehicle speed and Vy = side slip speed.

Furthermore, the cornering forces Cf Cr a response like a first order time lag, however neglect this delay the side leaders will like written as follows:

(Formula 6):

• Cf = Kfαf, Cr = Krαr, with Kf, Kr = lateral force coefficient of the front or rear wheel; and αf, αr = side slip angle of the front and rear wheels.

mit αf, αr = Lenkwinkel des Vorder- bzw. des Hinterrads und
n = Lenkgetriebeverhältnis.On the other hand, if the idea of an equivalent lateral force coefficient is introduced, taking into account the effects of roll of the vehicle or the suspension, the side slip angles αf, αr are written as follows: (Formula 7):

with αf, αr = steering angle of the front or rear wheel and
n = steering gear ratio.

Equations (3) to (7) above are fundamental equations of motion.

Various parameters are estimated by expressing these equations as a state variable and applying a parameter adjustment method to the theory of adaptive control. The lateral force coefficient is obtained from the parameters so estimated. With regard to the parameters of an actually built vehicle, there is a vehicle weight, a yaw moment of inertia and the like. When developing the theory, these vehicle parameters are assumed to be constant, and only the lateral force coefficient is assumed to be variable. The lateral force coefficient of the tire is variable in accordance with a non-linearity of the lateral force against the slip angle, an effect of the coefficient of friction of the road, an effect of the weight shift, and the like. If a is a parameter estimated by the change in the yaw angle γ and b is a parameter estimated on the basis of the front wheel steering angle δf, the lateral force coefficients of the front and rear wheels Kf, Kr are written as follows, for example:

(Formula 8):

• Kf = b · Iz · n / 2Lf Kr = (a · Iz + LfKf) / Lr

The lateral force coefficient of the front and the rear wheel Kf, Kr in the nonlinear range are estimated by Substitution of the vehicle speed V, the steering angle δf and the Yaw angle γ in the aforementioned Equations. Moreover by comparing the lateral force coefficients Kf, Kr with those on a road with a high coefficient of friction for everyone front and rear tires, for example the one below Way, a coefficient of friction μ the Lane calculated, and also based on the calculated Coefficient of friction μ an esteemed one Coefficient of friction E in the nonlinear range is determined with high accuracy.

(Formula 9)

• μf = Kf / KfO μr = Kr / KfO, with μf, μr = coefficient of friction of the front or rear wheel; Kf, Kr = estimated lateral force coefficient of the front or rear wheel; and KfO, KrO = equivalent lateral force coefficient of the front or rear wheel on the road with a high coefficient of friction. Here, the equivalent side force coefficients KfO, KrO are friction values given by correcting the tire characteristic, which is believed to produce cornering force proportional to the slip angle, in a range where a tire slip angle due to the characteristics of the Vehicle suspension is very small, among other things.

The above equations are to be understood as follows:
If the vehicle travels on a road with a high coefficient of friction with full tire grip and both the front and rear wheels are in the linear range of the tire characteristics, the estimated lateral force coefficients Kf, Kr can be regarded as equal to the equivalent lateral force coefficients KfO and KrO, and As a result, the coefficient of friction μs is estimated at 1.0. When the vehicle pushes outward or drifts, the slip angle of the front wheel becomes very large, and therefore, as in FIG 4 is shown, it is estimated that the estimated lateral force coefficient Kf = front wheel cornering force / front wheel slip angle becomes extremely small. If the vehicle or a wheel spins, the estimated lateral force coefficient Kr = rear wheel cornering force / rear wheel slip angle is also extremely small. To solve this problem to avoid, the larger of the estimated friction values for the front and rear wheels is determined as an estimated friction value "E" of the road surface.

Next, the case is described that the Torque is shared between the front and rear wheels.

α sollte in diesem Fall 0 ≤ α ≤ 1 sein.The equation of motion of a vehicle can be analyzed by extending it to the non-linear range using the vehicle speed V , the yaw angle γ of the input torque Ti, the target stability factor At, the estimated coefficient of friction E of the road, etc. The torque distribution ratio α between the front and Rear wheels are calculated according to the following equations of motion of the vehicle. (Formula 10):

In this case, α should be 0 ≤ α ≤ 1.

(Formel 12):

(Formel 13):

(Formel 14):

(Formel 15):

mit Gx' geschätzte Längsbeschleunigung; Gy' = geschätzte Querbeschleunigung; W = Fahrzeuggewicht; θ = Höhe des Schwerpunkts; L = Radstand; Lf = Entfernung zwischen dem Schwerpunkt und dem Vorderrad; Lr = Entfernung zwischen dem Schwerpunkt und dem Hinterrad; KfO, KrO = äquivalente Seitenkraftbeiwerte der Vorder- bzw. der Hinterräder im linearen Bereich; Kfc, Krc = Gewichtsabhängigkeit des Seitenkraftbeiwerts, der durch die Aufstandslast einem partiellen Differential unterliegt; Gt = Achsuntersetzung; Rt = Reifendurchmesser; Ti = Eingangsdrehkraft; At = Soll-Stabilitätsfaktor; At0 = Referenz-Sollstabilitätsfaktor (eine vorbestimmte Konstante, auf schwache Untersteuerung eingestellt); δf = Vorderrad-Lenkwinkel; Gγ = Gierwinkelzunahme; Δγ = Differenz zwischen dem Ist-Gierwinkel und dem Soll-Gierwinkel; und V = Fahrzeuggeschwindigkeit.If α> 1, α is left as 1, and if α <0, α is left as 0. Gx '= (TiGt / Rt) / (W / g) Gy '= V · γ (Formula 11):

(Formula 12):

(Formula 13):

(Formula 14):

(Formula 15):

longitudinal acceleration estimated with Gx '; Gy '= estimated lateral acceleration; W = vehicle weight; θ = height of the center of gravity; L = wheelbase; Lf = distance between the center of gravity and the front wheel; Lr = distance between the center of gravity and the rear wheel; KfO, KrO = equivalent lateral force coefficients of the front and rear wheels in the linear range; Kfc, Krc = weight dependence of the lateral force coefficient, which is subject to a partial differential due to the riot load; Gt = reduction gear; Rt = tire diameter; Ti = input torque; At = target stability factor; At0 = reference target stability factor (a predetermined constant set to weak understeer); δf = front wheel steering angle; Gγ = yaw angle increase; Δγ = difference between the actual yaw angle and the target yaw angle; and V = vehicle speed.

Based on the above equations, the control system according to 1 described.

Different data, one from a steering angle sensor 42 detected steering angle δf, one from a vehicle speed sensor 43 detected vehicle speed V , one from a yaw rate sensor 44 yaw angle γ detected by a speed sensor 45 detected engine speed N , one from an accelerator pedal angle sensor 46 detected accelerator pedal angle ϕ and one from a transmission position sensor 47 detected gear position P are in the control unit 50 entered.

In a friction estimation facility 51 the lateral force coefficients Kf, Kr of the front and rear wheels are estimated on the basis of the entered data, namely the steering angle δf, the vehicle speed V and the actual yaw angle γ, according to the aforementioned theory of adaptive control. The coefficient of friction of the road surface is calculated from the ratio of the estimated lateral force coefficients Kf, Kr to the equivalent lateral force coefficients KfO, KrO on the road surface with a high coefficient of friction (μ = 1.0). In order to avoid difficulties, for example, that the estimated lateral force coefficients of the front wheels become extremely small when the front wheels slide outwards, i.e. when the vehicle would not turn even when the steering wheel is turned, or that the estimated lateral force coefficients of the rear wheels become too small when the vehicle is rotated, a larger coefficient of friction than the estimated coefficient of friction E of the road surface is selected from the friction values of the front and rear wheels.

In the target yaw angle determination device 52 The target yaw angle γt is determined on the basis of the entered data, specifically the steering angle δf and the vehicle speed V. The target yaw angle δt and the actual yaw angle γ are in a target steering characteristic determination device 53 entered, in which the target stability factor At of the steering characteristic is determined and corrected according to the difference between the two yaw angles γt and γ. The stability factor is determined so that it has a slightly weak understeer characteristic in average vehicles. Therefore, when the vehicle spins or pushes outward, the stability factor At is set numerically according to the change in the actual yaw angle γ.

On the other hand, in the input torque estimator 54 an engine output Te is estimated from the input data, namely the engine speed N and the accelerator opening angle ϕ, and an input torque Ti of the intermediate differential is calculated by multiplying the estimated engine output Te by a gear ratio g in the gear position P.

These data, that is, vehicle speed V, actual yaw angle γ, input torque Ti, target stability factor At and estimated coefficient of friction E, are used by a torque distribution ratio computing device 55 in which the torque distribution ratio α between the front and rear wheels is calculated using the above equations. The torque distribution ratio α and the input torque Ti are input to a differential limiting torque calculator 56 in which an intermediate differential limiting torque Tc is calculated according to the following equation: Tc = (α - Di) Ti, with Di = reference torque distribution ratio, which is due to the combination of planet gears of the intermediate differential 20 is determined as already described. In this embodiment, the weight distribution between the front and rear wheels is directed towards the rear wheels. If the weight distribution is directed towards the front wheels, the above equation is rewritten as follows: Tc = (Di - α) · Ti.

In these equations, if the calculated intermediate differential limiting torque Tc is negative, Tc = 0. The torque Tc thus calculated is calculated in the computing device 56 converted to a torque signal for the intermediate differential limiting torque, and the torque signal becomes a duty ratio converter 57 supplied, in which it is converted into a certain duty cycle D, and this duty cycle is applied to the solenoid valve 40 issued.

Next is the operation of this embodiment described.

First, the output power of the engine 1 through the clutch 2 passed into the gearbox, and the converted power becomes the first sun gear 21 of the intermediate differential 20 fed. As described above, since the reference torque distribution ratio et is set to be directed toward the rear wheels, the driving force is applied to the planet carrier 24 and the second sun gear 22 given with this torque distribution ratio. If the central clutch 27 is disengaged, the driving force is transmitted with this distribution ratio et to the front and rear wheels.

As a result, the vehicle exhibits driving behavior like a vehicle with a front engine / rear-wheel drive. Because the intermediate differential 20 is free, the vehicle can take turns without hindrance, while at the same time absorbing the rotational difference between the front and rear wheels. If doing so the on-time signal from the control unit 50 to the solenoid valve 40 is delivered by the hydraulic control device 32 generates the differential limiting torque Tc. The torque Tc bypasses the second sun gear and the planet carrier 24 and is transmitted to the front wheels. As a result, a larger torque is distributed to the front wheels than to the rear wheels, so that the torque force distribution toward the front wheels is obtained.

While the vehicle is running, signals relating to the steering angle δf, the vehicle speed V and the actual yaw angle γ are fed into the control unit 50 entered, and the vehicle behavior is constantly monitored. When the vehicle is running on a road whose road surface has a large coefficient of friction, the actual yaw angle γ approximately coincides with the target yaw angle γt based on the steering angle δf and the vehicle speed V in the target yaw angle determining device 52 is determined. As a result, the stability factor At is predetermined with weak understeer, and the steering characteristic of the vehicle is constantly kept with weak understeer.

The from the road surface friction estimation device 51 calculated estimated friction values E become the torque distribution ratio calculator 55 transmitted, in which the torque distribution ratio α is calculated on the basis of the following data: this calculated coefficient of friction E, the vehicle speed V, the actual yaw angle γ, the stability factor At and the input torque Ti.

When the vehicle is traveling straight ahead, the torque distribution ratio α is primarily based on the Input torque Ti and the estimated longitudinal acceleration Gx 'determined.

When the vehicle is cornering the torque distribution ratio α primarily on the Basis of the vehicle speed V and the estimated lateral acceleration Gy 'determined that determined by the actual yaw angle γ is. Since the actual yaw angle γ is one Feedback control the control system is subject to malfunctions or control errors unaffected.

First, when the vehicle turns or turns on the low-friction road and the torque is distributed more to the rear wheels than to the front wheels, the cornering force of the tire on the rear wheel side is reduced due to excessive rear wheel traction, and as a result Rear wheels slip in the transverse direction. Then when the tire grip finally exceeds a limit value and the vehicle starts to spin, the road surface friction estimator is used 51 the lateral force coefficients Kf, Kr of the front and rear wheels are estimated on the basis of the steering angle δf, the vehicle speed V and the actual yaw angle γ in accordance with the vehicle behavior. For each of the front and rear wheels, the coefficient of friction of the road is calculated by comparing the coefficient of friction with that of the road with the highest coefficient of friction, and the highest among the coefficients of friction thus calculated is selected. This highest coefficient of friction is the estimated coefficient of friction E.

In the target steering characteristic determination device 53 Further, the target stability factor At is determined in accordance with the above-described equation (15) on the basis of the yaw rate increase G γ by the yaw rate increase determining means 58 is determined, and the difference Δγ between the actual yaw angle γ and the target yaw angle γt, that of the target yaw angle determining device 52 is determined. For example, if the actual yaw angle becomes greater than the target yaw angle as a result of the vehicle spinning, the target stability factor At becomes greater than the reference target stability factor AtO, ie the target stability factor At is corrected in the direction of a stronger understeer. Then in the torque distribution ratio calculator 55 the torque distribution ratio α is calculated and controlled by being directed toward the front wheels, and as a result, the cornering force of the rear wheels becomes larger so that the vehicle does not spin.

The feedback control is thus determined by the target stability factor At so done that the actual yaw angle matches the target yaw angle, and thus the vehicle behavior is always in the state of a Great weak understeer held.

Next is the control of the Distribution of torque between left and right wheels is described. Here is an example of the control between left and right wheels the torque distribution control between the left and right Rear wheel explained.

The torque distribution control system this embodiment is based on the principle below.

If a rear differential differential torque Td during a fast cornering with the accelerator pedal depressed increases the braking force of the outer rear wheel larger than that of the inner rear wheel, and as a result creates the difference between these braking forces a moment M aimed to keep the vehicle straight ahead drive. It is known that this Torque M is effective to prevent the vehicle from being pulled inwards. On the other hand, the extent of Pushing inward the deviation of the actual yaw angle γ from that Target yaw angle γt judged by the driver of a vehicle according to the Vehicle speed V and the steering angle δf determined when driving through the curve. If the numerical extent of pulling in as a change of the stability factor identified pulling inwards by generating a yaw moment M can be prevented, so this change in the stability factor repealed. This means so that for Avoid pushing inwards Rear axle differential limiting torque Td can be determined in this way should that Yaw moment M is generated.

First, the target yaw rate γ becomes as follows certainly.

According to a two-wheel model 6 equations of motion are written as follows:

(Formula 16):

• mv (. + γ) = Cf + Cr I. = Lf · Cf - Lr · Cr with γ = actual yaw angle; β = vehicle slip angle; V = vehicle speed (V is constant); m = vehicle mass; I = yaw moment of inertia; Cf, Cr = lateral force coefficient of the front and rear wheel; Lf, Lr = distance between the center of gravity and an axle of the front or rear wheel.

The relationship between the lateral force coefficient and the tire slip angle in the linear region is written as Cf = 2Kf · αf, Cr = 2Kr · αr, where Kf, Kr is an equivalent lateral force coefficient of the Vor der- or the rear wheel and αf, αr is a slip angle of the tire of the front or rear wheel.

Introducing the above relationship in the equations (formula 16) the equations of motion become further written as follows:

(Formula 17):

• mv. + 2 (Kf + Kr) β + {mv + 2 (LfKf - LrKr) / V} γ = 2Kfδf + 2Krδr 2 (LfKf - LrKr) β + I. + {2 (Lf 2 Kf + Lr 2 Kr) / V} γ = 2LfKfδf - 2LrKr .δr

Based on the above fundamental The target yaw angle γt is obtained for vehicle motion equations.

Next, how the yaw moment M and the rear differential differential torque Td can be calculated.

When the rear differential differential torque Td in equations of motion of the two-wheel vehicle model of 6 is introduced, they are written as follows:

(Formula 18):

• mv (. + γ) = Cf + Cr (1) I. = LfCf - LrCr - M (2) with γ = yaw angle (variable); β = vehicle slip angle (variable); m = vehicle mass; V = vehicle speed; Cf, Cr = lateral force coefficient of the front and rear wheel; I = yaw moment of inertia; Lf, Lr = distance between the center of gravity and an axis of the front or rear wheel; M = torque generated by the rear differential differential torque.

The cornering power of the front and of the rear wheel is written as follows:

(Formula 19):

• Cf = 2Kfαf Cr = 2Krαr (3) with Cf, Cr = cornering force of the front or rear wheel; Kf, Kr = lateral force coefficient of the front or rear wheel; αf, αr = a slip angle of the tire of the front or rear wheel.

If the steering angle δf or δr in the Tire slip angle αf or αr substituted be leads the substitution of equation (3) into equations (1) and (2) to the following equations:

(Formula 20):

• mv. +2 (Kf + Kr) β + {mv + 2 (LfKf - LrKr) / V} γ = 2Kfδf + 2Krδr (4) 2 (LfKf - LrKr) β + I. + {2 (Lf 2 kf + Lr 2 Kr) / V} γ + M = 2LfKfδf - 2LrKrδr (5)

Next are the characteristics of the vehicle, when it runs on a fixed circle. In this case, both the slip angle β of the vehicle as well as the yaw angle γ constant, and Deviations from this are left zero his. Equations (4) and (5) are written as follows:

(Formula 21):

• 2 (Kf + Kr) β + {mv + 2 (LfKf - LrKr) / v} γ = 2Kfδf (6) 2 (LfKf - LrKr) β + {2 (Lf 2 Kf + Lr 2 Kr) / v} γ + M = 2LfKfδf (7) the steering angle δf of the rear wheel should be zero.

Equation (7) is as follows Equation transferred:

(Formula 22):

• 2 (LfKf - LrKr) β + {2 (Lf 2 Kf + Lr 2 Kr) / v + M / γ} / γ = 2LfKfδf (8)

mit L = Radstand (Lf + Lr); γ auf der rechten Seite = vorher erhaltener Gierwinkel.The solution of γ is given by equations (6) and (8) as follows: (Formula 23):

with L = wheelbase (Lf + Lr); γ on the right side = yaw angle previously obtained.

In order for equation (9) to have a physical meaning, the following condition must be met: (Formula 24):

mit A = Stabilitätsfaktor in einem Fall, in dem die Hinterachsdifferential-Drehkraftbegrenzungssteuerung frei ist.With the introduction of a stability factor A ', which is applied to the vehicle with a rear axle differential limitation control when expanded, the following equation is obtained: (Formula 25):

with A = stability factor in a case where the rear axle differential torque control is free.

Therefore, if the yaw angle γ is increased by pulling inwards (Δγ> 0), the deviation ΔA of the stability factor is written as follows: (Formula 26):

In the above equation, Gγ denotes an increase in the yaw angle of the steering angle δt of the front wheel, and the increase in the yaw angle is written as follows: (Formula 27):

As a result, the moment M required to cancel the inward pull is written as follows: (Formula 28):

Furthermore, the rear differential differential torque Given Td as follows:

(Formula 29):

• Td = (M / d) · R with R = tire diameter and d = tread.

With reference to 5 the operation of the torque distribution control system will now be described.

Signals of the yaw rate γ from the yaw rate sensor 44 of the steering angle δf is detected by the steering angle sensor 42 and the vehicle speed V by the vehicle speed sensor 43 is detected, are in a control unit 70 entered. In the control unit 70 is a yaw rate increasing determiner 71 provided in which a yaw angle increase Gγ of the predetermined steering angle δf of the front wheel is determined either from the aforementioned equations or by reading from a table. The vehicle speed V and the steering angle δf are converted into a target yaw angle computing device 72 entered, in which a target yaw angle gt, which corresponds to the driving condition on the road with a high coefficient of friction, is calculated on the basis of the aforementioned equations of motion. The calculated target yaw angle γt and the actual yaw angle g are in a differential computing device 62 entered, in which the difference Δγ (Δγ _{= γ - γ} t with Δγ> 0) is calculated. Thus, pulling inward is detected by an increase in the actual yaw angle γ, and also the amount of pulling inward is obtained from the deviation Δγ.

The yaw rate increase Gγ and the yaw rate deviation Δγ, which correspond to the amount of pulling inward, are converted into a yaw moment calculator 74 entered. In this, a deviation of the stability factor is first obtained as ΔA using the yaw increase Gγ and the yaw deviation Δγ. Since the stability factor is predefined on the slightly understeering side, if a yaw angle deviation Δγ is generated by pulling inwards, the deviation ΔA of the stability factor becomes a negative value (i.e. on the oversteer side) corresponding to the yaw angle deviation Δγ. Finally, a yaw moment M is calculated on the basis of the calculated deviation ΔA, which is required to cancel the deviation ΔA.

The yaw moment M is in the rear axle differential limiting torque calculator 75 entered in which a rear axle differential limiting torque Td is calculated. This torque signal Td is then in the duty cycle converter 76 converted to a duty cycle D, and then the duty cycle signal D to the solenoid valve 40 ' issued.

That of the intermediate differential 20 and the central coupling 27 distributed torque is applied to the rear axle differential 11 transfer. If the rear axle coupling 28 is disengaged, distributes the rear axle differential 11 the driving force evenly on the left rear wheel 13L and the right rear wheel 13R , Furthermore, if the accelerator pedal is released in this case, the braking force is also distributed evenly. If the rear axle coupling 28 from the hydraulic control device 32 ' is engaged in the rear axle coupling 28 generates a differential limiting torque Td, and the torque distribution between the left and right rear wheels 13L and 13L is changed by the differential limiting operation. Thus, when the driving force is applied, the torque is transmitted from the fast wheel to the slow wheel (the wheel with grip) in accordance with the rear axle differential limiting torque Td. On the other hand, if the rotational speed of the outer wheel is greater than that of the inner wheel while the vehicle is turning through a curve when the accelerator pedal is released, the braking force is distributed more to the outer wheel than to the inner wheel in accordance with the rear axle differential limiting torque Td.

In the operation of the vehicle are in the control unit 70 Signals of the steering angle δt, the vehicle speed V and the yaw angle γ are input, and the vehicle behavior is continuously monitored. If the vehicle behavior does not change while the vehicle is driving straight ahead or cornering, the target yaw angle γt, that of the steering angle δf and the vehicle speed V in the target yaw angle computing device, is correct 72 was calculated, coincides with the actual yaw angle γ, and therefore the Stability factor is not. Thus, the rear differential differential torque Td remains zero.

On the other hand, when the vehicle comes into the inward-pulling state, that is, when the vehicle abruptly turns inward while driving through a curve at high speed with the accelerator pedal released, the actual yaw angle γ becomes larger. The deviation calculator then calculates 73 the deviation Δγ of the actual yaw angle γ from the target yaw angle γt, and the amount of pulling inward is detected. In the yaw moment calculator 74 this deviation Δγ is then converted into the deviation ΔA of the stability factor, and then the yaw moment M for canceling this deviation ΔA is calculated therein.

Then in the computing device 75 for the rear axle differential limiting torque, the rear axle differential limiting torque Td is calculated according to the calculated yaw moment M, and this torque Td is applied to the rear axle clutch 28 applied. Accordingly, when the vehicle rotates 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 rear inner wheel, and as a result, the torque M that releases the inward pushing is generated, so that this phenomenon is prevented. Further, since feedback control is performed in the control system of this embodiment so that the actual yaw rate γ matches the target yaw rate γt, the vehicle never goes to the side of an unfavorable strong understeer, and the phenomenon of pulling inward can be secured be avoided. Furthermore, since the control system is constructed so that the yaw angle deviation Δγ is converted into the deviation ΔA of the stability factor, the steering characteristic maintains a weak understeer for which it was originally designed.

In this embodiment, the torque distribution control system became an example of controlling the torque distribution between described the left and right rear wheels. The basic features of the tax system are, however, also for a tax system for the division the torque between the left and right front wheels applicable.

How 4 shows, the lateral force coefficients Kf, Kr, which correspond to pushing the vehicle outwards or spinning in the limit area, are exactly determined at any moment by reducing or increasing the previously obtained lateral force coefficients by a predetermined increment in accordance with Table 1.

Thus, estimated lateral force coefficients Kf, Kr of the front and rear wheels are in the coefficient of friction determination device 64 entered, and the coefficient of friction of the front and rear wheels is estimated in each case by comparing the estimated lateral force coefficients with those of the road with a high coefficient of friction. The estimated coefficient of friction E is a larger coefficient of friction among the above estimated coefficients of friction.

In summary, the torque distribution control system for safe and comfortable driving in all road conditions or in provides a limit of vehicle behavior by driving torque right on the wheels is distributed.

The control mechanism for torque distribution on the front and rear wheels comprises a friction estimation device to appreciate a coefficient of friction of the road surface due to the lateral force coefficients of the Front and rear wheels, a target yaw angle determining device for determining a target yaw angle according to proviso the driving condition of the vehicle on the road with a high coefficient of friction, based on a steering angle and a vehicle speed, a target steering characteristic determination device to determine a target stability factor according to the Difference between the target and actual yaw angles, an input torque estimator to appreciate an input torque of the intermediate differential, a torque distribution ratio calculator for calculating a torque distribution ratio between the front and rear wheels from the vehicle's equations of motion, which are based on a non-linear Range can be expanded based on vehicle speed, the actual yaw angle, the input torque, the estimated coefficient of friction the road surface and the target stability factor, and computing means for computing an intermediate differential limiting torque based on the torque distribution ratio between the front and rear wheels and the input torque.

In the torque distribution control mechanism thus constructed between the front and rear wheels is the torque when the vehicle is on a high friction road moves, between the front and rear wheels correctly distributed, according to the driving conditions of the Vehicle such as driving straight ahead and cornering, so that handiness of the vehicle is excellent.

If, on the other hand, the vehicle is on a Lane with a low coefficient of friction, can by a coefficient of friction estimated with high accuracy and the torque between the front and rear wheels is correct distributed, a spinning or pushing the vehicle outwards be prevented.

The torque distribution mechanism between the left and right rear wheels includes a target yaw angle calculator for calculating a target yaw angle from vehicle motion equations based on the steering angle and the vehicle speed, a deviation calculator for calculating a deviation of the actual yaw angle from the target yaw angle according to the extent pulling inward, a yaw rate increase determining means for determining a yaw rate increase based on a predetermined table in which the vehicle speed is included as parameters, a yaw moment calculating means for calculating a yaw moment required to cancel the deviation of the stability factor, the was calculated from the above deviation of the actual yaw angle and the above yaw angle increase, and a computing device for calculating a rear axle differential limiting torque in accordance with the above yaw moment.

Because in the torque distribution mechanism thus constructed between the left and right rear wheels the extent of the inward pushing exactly can be detected, a rear axle differential limiting torque, required to release the inward push is calculated become. The calculated rear axle differential limiting torque generates a yaw moment so that the Phenomenon of Pushing the Vehicle can be prevented.

Claims (2)

Procedure for determining the road coefficient of friction between road and tire for a control system of a motor vehicle with the following steps: - To calculate the lateral force coefficient of at least one front and rear wheel by applying the vehicle motion equations known per se based on the determination of the steering angle, the vehicle speed and the yaw angle, - Determine the road friction by comparing the calculated lateral force coefficients with those on a road with a high coefficient of friction for the wheels. Use of the method according to claim 1 for control the driving force distribution between front and rear wheels and / or the left or right wheels an axle of a motor vehicle with the following steps: - Determine a coefficient of friction of the road surface by calculating current lateral force coefficients of the front and rear wheel using the cornering forces of the Tires determined using the vehicle motion equation the larger being the certain coefficients of friction for The front and rear wheel is determined as the coefficient of friction of the road, continue from the steering angle and vehicle speed ongoing calculation of a target yaw angle and a target lateral acceleration taking into account the previous lateral force coefficients is made - ongoing Determination of the actual yaw angle and the actual lateral acceleration, the difference between debit and Actual values of yaw angle and lateral acceleration is determined and depending on this current, current lateral force coefficients are determined by the difference become, - ongoing Determination of a target stability factor for the Steering characteristic from the difference between target and actual yaw angle, - ongoing Determination of an input torque taking engine output power into account and gear ratio, - recourse on the data from the previous steps to determine the Driving force distribution α between Front and rear wheels, whereby the distribution ratio α is changed so that the current Target stability factor to a reference stability factor is traceable, so that at Cornering a match of target and actual yaw angle and freedom of slip when driving straight ahead the drive wheels and maintaining or returning to the target stability factor is achieved and that for Distribution of driving force between the left and right wheels of the calculated target yaw angle with the measured actual yaw angle Finding a difference is compared, with an increase the yaw angle of the front wheel depending on the driving speed is determined, and being continued based on the difference the yaw angle and the determined, speed-dependent increase the yaw angle of the front wheel a yaw moment to cancel the Deviation and reaching the stability factor and that this Yaw moment for determining a rear axle differential limiting force is used.