JP2712786B2 - Braking force control device - Google Patents

Description

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a braking force control device that can improve the steering stability of a vehicle during braking.

[Conventional technology]

As a conventional braking force control device, for example,
As described in Japanese Patent Publication No. 264, there is an apparatus that controls the vehicle yaw characteristic by a differential pressure between left and right brakes. Specifically, when braking is performed with the driver's steering angle being equal to or larger than a predetermined value, control is performed such that the pressure increase timing of the turning outer wheel is delayed to improve the turning performance during braking.

[Problems to be solved by the invention]

However, the conventional braking force control device does not consider that the yaw rate generated by the front wheel steering and the difference between the left and right braking forces depends on the vehicle speed, so that the yaw rate cannot be controlled to an appropriate value and the generated yaw rate cannot be controlled. There is an unsolved problem that the transient characteristics cannot be improved.

In view of the above, the present invention has been made by focusing on the unsolved problems of the conventional example, and can control the yaw rate to an appropriate value and improve the transient characteristics of the generated yaw rate. It is an object to provide a braking force control device.

[Means for solving the problem]

In order to achieve the above object, a braking force control device according to claim (1) includes a steering state detecting means for detecting a steering state of a vehicle, as shown in a basic configuration diagram of FIG. Speed detection means for detecting the front-rear direction speed of the vehicle, target yaw rate setting means for setting a target yaw rate of the vehicle based on the detection values of the steering state detection means and the speed detection means, and at least one of the front wheels and the rear wheels. Braking force control means capable of independently controlling the braking forces of the left and right braking means, and a front wheel and a rear wheel necessary for realizing the target yaw rate set by the target yaw rate setting means in the vehicle to be controlled. At least one of the left and right target braking forces is determined in advance by at least the target yaw rate, the differential value of the target yaw rate, the detection value of the steering state detecting means, and vehicle specifications and a motion equation in advance. A target braking force calculation unit that calculates by a calculation based on the determined vehicle model, wherein the braking force control unit matches a braking force of the braking unit with a target braking force calculated by the target braking force calculation unit. So that it is controlled.

Further, as shown in the basic configuration diagram of FIG. 1 (b), the braking force control device according to claim (2) detects a steering state detecting means for detecting a steering state of the vehicle, and detects a longitudinal speed of the vehicle. Speed detection means, braking pressure detection means for detecting the braking pressure of the four-wheel braking means, target yaw rate setting means for setting a target yaw rate of the vehicle based on the detection values of the steering state detection means and the speed detection means. A braking force control means capable of independently controlling the braking forces of the left and right braking means disposed on at least one of the front wheel and the rear wheel; and a target yaw rate set by the target yaw rate setting means in a vehicle to be controlled. Target braking force calculating means for calculating the left and right target braking forces of at least one of the front wheels and the rear wheels required for realization by calculation based on a vehicle model set in advance by vehicle specifications and equations of motion Has, the target braking force calculation means is changed in accordance with the value corresponding to the cornering power of the wheels of the vehicle model on the detected value of the braking pressure detecting means,
The braking force control means controls the braking force of the braking means so as to match the target braking force calculated by the target braking force calculation means.

Further, in the braking pressure control device according to claim (3), the target braking force calculating means determines the cornering power of the wheel in the vehicle model based on the detection of the braking pressure detecting means and the radius of the friction circle accompanying the load movement. It is configured to change the corresponding value.

[Action]

In the braking force control device according to the first aspect, the target yaw rate _r is calculated by the target yaw rate setting means based on the detected value of the steering state of the vehicle, for example, the detected steering angle and the speed in the front-rear direction of the vehicle, for example, the vehicle speed. By calculating the target yaw rate _{r based} on the vehicle speed and the steering angle in this manner, it is possible to obtain an optimum target yaw rate having a dependency on the vehicle speed. Then, to match the yaw rate occurring in actual vehicle and the target yaw rate _r, the target braking force calculating means, at least the target yaw rate _r, a differential value _r and the steering state detection value of the target yaw rate _r, vehicle specifications and exercise Calculation is performed based on the vehicle model set by the equation to calculate a target braking force that causes a braking force difference between the left and right braking force control means. By supplying this target braking force to the braking force control means and controlling the braking force generated by the braking means to match the target braking force,
The vehicle yaw rate is controlled to an optimum value to improve the steering stability of the vehicle and to improve transient yaw rate characteristics.

Also, in the braking force control device according to claim (2), the braking pressure of each wheel of the vehicle is detected by the braking pressure detecting means and detected by the target braking force calculating means. By calculating a value corresponding to the cornering power in the vehicle model based on the detected braking pressure, an accurate cornering power corresponding to the braking force generated by the vehicle is calculated, and the target braking force is calculated using the cornering power. , The deviation between the yaw rate characteristics of the vehicle model and the actual vehicle is corrected, and the steering stability of the vehicle is further improved, and the transient yaw rate characteristics are further improved.

Further, in the braking force control device according to claim (3), the value corresponding to the cornering power is calculated based on the detected braking pressure value and the change in the radius of the friction circle accompanying the movement of the load during braking. A target braking force can be calculated, and the deviation between the vehicle model and the actual yaw rate characteristic can be corrected more accurately.

〔Example〕

 Hereinafter, embodiments of the present invention will be described with reference to the drawings.

FIG. 2 is a hydraulic system diagram showing a first embodiment of the present invention.

In the figure, 1FL and 1FR are wheel cylinders as braking means attached to front wheels, 1RL and 1RR are wheel cylinders as braking means attached to rear wheels, and brakes supplied to these wheel cylinders 1FL to 1RR. The hydraulic pressure is controlled by the actuator 2.

This actuator 2 is a wheel cylinder on the front wheel side.
Equipped with 3 port 3 position electromagnetic directional control valves 3FL and 3FR for individually controlling the cylinder pressures of 1FL and 1RR, and 3 port 3 position electromagnetic directional control valve 3R for simultaneously controlling the rear wheel cylinders 1RL and 1RR. I have.

The P ports of the solenoid-operated directional control valves 3FL and 3FR are connected to one system of a two-system master cylinder 5 connected to the brake pedal 4, and the solenoid-operated directional control valves 3FL and 3FR are connected.
Port A is individually connected to the wheel cylinders 1FL and 1FR, and port B is connected to one system of the master cylinder 5 via a hydraulic pump 7F that is driven to rotate by an electric motor (not shown).

The A port of the electromagnetic directional control valve 3R is connected to the wheel cylinders 1RL and 1RR, and the B port is connected to the other system of the master cylinder 5 via a hydraulic pump 7R driven to rotate by an electric motor (not shown). Have been.

Further, an accumulator 8F is connected to a pipeline between the P ports of the electromagnetic directional valves 3FL and 3RR and the hydraulic pump 7F,
Reservoir tank in the line between B port and hydraulic pump 7F
9F is connected. Similarly, an accumulator 8R is connected to a pipe between the P port of the electromagnetic directional control valve 3R and the hydraulic pump 7R, and a reservoir tank 9R is connected to a pipe between the B port and the hydraulic pump 7R. Have been.

Here, each of the electromagnetic directional control valves 3FL to 3R directly connects the master cylinder 5 and the wheel cylinders 1FL to 1RR at the first switching position of the normal position, and
In a pressure increasing state in which the brake fluid pressure of 1FL to 1RR is set to a value corresponding to the master cylinder 5, the wheel cylinders 1FL to 1RR are disconnected from the master cylinder 5 and the hydraulic pumps 7F and 7R at the second switching position. By holding the brake fluid pressure of the wheel cylinders 1FL to 1RR in a holding state, and connecting the wheel cylinders 1FL to 1RR and the master cylinder 5 at the third switching position via a hydraulic pump, the wheel cylinders 1FL to 1RR are connected. The hydraulic pressure in 1RR is reduced to return to the master cylinder 5 side, and these switching positions are switched and controlled by three stages of current values supplied from a braking pressure control device 16 described later.

On the other hand, in the vehicle, the steering angle of the steering wheel 10 is detected, a zero voltage when the steering wheel 10 is in the neutral position, a negative voltage according to the steering angle when turning right from the neutral position, and A steering angle sensor 11 is provided as a steering state detecting means for outputting a steering angle detection value θ which becomes a positive voltage corresponding to the steering angle when the vehicle turns left from the neutral position, and a vehicle speed detection corresponding to the vehicle speed is provided. vehicle speed sensor 12 is attached as a speed detecting means for outputting a value V _X, also with a brake switch 13 for detecting a depression state of the brake pedal 4 is mounted, the wheel cylinders 1F
L, 1FR, 1RL, 1RR and the pressure detection value P _FL corresponding to the cylinder pressure of the master cylinder 5, P _FR, the pressure sensor 14FL that outputs P _R and P _MC, 14FR, 14R and 14MC are attached, the sensors Is input to the braking pressure control device 16.

As shown in FIG. 3, the braking pressure control device 16
11, 12, 13 and a microcomputer 19 that the detected values of 14FL~14MC is input, the electromagnetic direction control signal CS _FL outputted from the microcomputer 19, the CS _FR and CS _R is described above is input separately It is provided with floating type constant current circuits 20FL, 20FR and 20R for driving solenoids of the switching valves 3FL, 3FR and 3R.

The microcomputer 19 includes at least an input interface circuit 19a having an A / D conversion function, an output interface circuit 19b having a D / A conversion function, an arithmetic processing device 19c, and a storage device 19d.
Steering angle detection value θ from vehicle, vehicle speed detection value from vehicle speed sensor 12
Master cylinder pressure detection value from the V _X and the pressure sensor 14MC
Together executes processing of FIG. 5 for calculating a target wheel cylinder pressure P ^* _FR and P ^* _FL of the target braking force of the front wheel left and right on the basis of the P _MC, these target wheel cylinder pressure P ^* _FR and P ^* _FL a pressure sensor 14FR, 14FL and the cylinder pressure detection value P _FR of 14MC, P _FL and P _MC and based on the fifth diagram of processing executed by the electromagnetic directional control valve of the actuator 2 3FL, control signal CS for controlling the 3FR Outputs _FL , CS _FR and solenoid directional control valve 3R
For outputs a control signal CS _R always zero.

 Next, the operation of the above embodiment will be described.

First, the control principle of the first embodiment will be described. Assuming that the motion of the vehicle is two degrees of freedom in yaw and lateral directions as shown in FIG. 4, the equations of motion are expressed by the following equations (1) and (1). 2) It can be expressed by the following equation.

I _Z · (t) = C _f · L _f- C _r · L _r + T _f · (B _FL (t)-B _FR
(T)) / 2 ............ ( 1) M · y (t) = 2 (C f + C r) -M · V X (t) · ............ (2) where, I _Z is the vehicle Yaw moment of inertia, is the yaw rate, L _f is the distance between the vehicle center of gravity and the front axle, L _r is the distance between the vehicle center of gravity and the rear axle, T _f is the front wheel tread, B _FL (t)
The left front wheel braking force, B _FR (t) is the right front wheel braking force, M is vehicle weight, V _y is the vehicle lateral velocity, _{y (t)} is the vehicle lateral acceleration, V _X is the vehicle longitudinal direction speed. Also, C _f and C
_r is the cornering force of the front and rear wheels,
It can be expressed by the following equations (3) and (4).

C _f = K _f ｛θ (t) / N− (V _y + L _f · (t)) / V
_x (t)｝ (3) C _r = −K _r (V _y −L _r · (t)) / V _x (t) (4) where θ (t) is steering Angle, N is the steering gear ratio, K _f
Is front wheel cornering power, and K _r is rear wheel cornering power.

Equations (3) and (4) are replaced by equations (1) and (2).
Yaw rate (t), lateral velocity V _y (t)
When considered as a differential equation, the following equations (5) and (6) can be expressed.

_{(T) = a 11 · (} t) + a 12 · V y (t) + b 1 · θ
(T) + b _PL · ΔB _f (t) (5) _y (t) = a ₂₁ · (t) + a ₂₂ · V _y (t) + b ₂ · θ
(T)... (6) where ΔB _f (t) = B _FL (t) −B _FR (t)... (7) a ₁₁ = −2 (K _f · L _f · L _{f + K r · L r · L} r) / (I z · V x) ............ (8) a 12 = -2 (K f · L f -K r · L r) / (I z · V x) _{............ (9) a 21 = -2} (K f · L f -K r · L r) / (M · V x) -V x ............ (10) a 22 = -2 (K f + K _r ) / (M · V _x ) (11) b ₁ = 2 · K _f · L _f / (I _z · N) ··· (12) b ₂ = 2 · K _f / ( M · N) (13) b _P1 = T _f / (2 · I _z ) (14)

From the above equations (5) and (6), the steering angle input θ (t)
The relation of the generated yaw rate ₁ (t) to the following can be expressed by the following equation (15) using the differential operator s.

The (15) of the transfer function X (s) is (primary) / in the form of (secondary), V _X is a steering angle input greater the larger theta (t)
It can be seen that the generated yaw rate ₁ (t) with respect to is oscillating, and the vehicle maneuverability and stability deteriorate. That is, the coefficient ｛− (a
₁₁ + a ₂₂ ) S} corresponds to the damping coefficient 制 御 of the control system. Therefore, the coefficient {− (a ₁₁ + a ₂₂ ) S} is obtained by adding a ₁₁ and a ₂₂ shown in the above equations (8) and (9). Substituting, these a ₁₁ , a
_{Since 22} is always a negative value, the ζ damping coefficient is a positive damping, and the damping extent vehicle longitudinal direction velocity V _X becomes large coefficient ζ will be close to zero. That, as the vehicle longitudinal direction velocity V _X becomes large, the damping coefficient of the control system ζ decreases, the yaw rate _{1 (t)} is oscillatory (hardly attenuated).

Therefore, for example, if the target yaw rate _r (t) is a first-order lag system without overshoot and undershoot with respect to the steering angle input θ (t) and the steady-state value is set equal to that of the normal vehicle, the target yaw rate _r (t) t) can be represented by the following equation (16). _{r (t) = H 0 ·} θ (t) / (1 + τS) ............ (16) where, H ₀ is the steady yaw rate gain, by using a stability factor A, is defined by the following expression (17) You.

H ₀ = V _x / ｛(1 + A · V _{x 2} ) · L · N) (17) where L is the wheel base, and the stability factor A is It is represented by

As described above, since the steady value is set to be equal to that of the normal vehicle in the above equation (16), the turning performance is not positively improved at the time of braking the vehicle as in the conventional example described above. when, by satisfying the predetermined yaw characteristics defined by the steering angle detection value θ and the vehicle speed detection value V _X irrespective during non-braking, it is possible to improve the steering stability of the vehicle.

Next, using the braking force difference ΔB _f (t) between the left and right front wheels, the generated yaw rate (t) of the vehicle is set to the target yaw rate ψ _r (t).
A method for matching the values will be described. By modifying the equation (16), the differential value _r (t) of the target yaw rate can be obtained by the following equation (19). _{r (t) = H 0 ·} θ (t) / τ- r (t) / τ ............ (19) steering angle input θ (t) and the front wheel left and right braking force difference ΔB _f (t) due to the occurrence yaw rate ( t) is the target yaw rate
Assuming that _r (t) matches, the respective differential values (t) and _r (t) also match. Therefore,
Assuming that _r (t) = (t), _r (t) = (t),
Further, the lateral velocity V _y (t) at the time when the above assumption is satisfied is represented by V _yr
By defining these as (t) and substituting these into the above equations (5) and (6), the following equations (20) and (21) can be obtained. _{r (t) = a 11 ·} r (t) + a 12 · V yr (t) + b 1 ·
θ (t) + b _P1 · ΔB _f (t) (20) _yr (t) = a ₂₁ · _r (t) + a ₂₂ · V _yr (t) + b ₂ ·
θ (t) (21) By modifying the above equation (20), the braking force difference ΔB _f (t) between the left and right front wheels can be obtained by the following equation (22).

The braking force difference ΔB _f (t) between the left and right front wheels obtained by the equation (22)
In order to generate the difference, the pressure difference between the wheel cylinder pressures on the left and right of the front wheel may be generated, and the relationship between the wheel cylinder pressure P and the braking force _Bf can be expressed by the following (23) It can be obtained by the formula.

_{B f = 2 · μ p ·} A p · r p · P / R = k p · P ............ (23) k p = 2 · μ p · A p · r p / R ............ (24 ) However, k _p is a proportionality constant between the wheel cylinder pressure and the braking force, mu _p brake pad and the disc rotor friction coefficient between, a _p is the wheel cylinder area, r _p is the disc rotor effective radius, R represents the tire radial It is.

Therefore, assuming that the target differential pressure of the wheel cylinder pressures of the front left and right wheels is ΔP (t), the target differential pressure ΔP (t) is ΔP (t) = ΔB _f (t) / k _p (25) ).

Then, the target differential pressure ΔP obtained by the above equation (25)
From (t) and the master cylinder pressure P _MC (t), target wheel cylinder pressures P _FL ^* (t) and P _FR ^* (t) for the front left and right wheels.
Is calculated according to the following equations (26) and (27).

_{^{P * FL (t) = P}} MC (t) (ΔP (t) ≧ 0) = P MC (t) + ΔP (t) (ΔP (t) <0 and _{P MC (t)> -ΔP (} t)) = 0 (ΔP (t) < 0 and _{P MC (t) ≦ -ΔP (} t)) ............ (26) P * FR (t) = P MC (t) (ΔP (t) <0) = _{P MC (t) -ΔP (t} ) (ΔP (t) ≧ 0 and _{P MC (t)> ΔP (} t)) = 0 (ΔP (t) ≧ 0 and _{P MC (t) ≦ ΔP (} t)) ………… (27) Therefore, the arithmetic processing unit of the microcomputer 19
At 19c, by executing the target wheel cylinder pressure calculation processing of FIG. 5 and the braking force control processing of FIG. 6, the braking force on the front left and right wheels is controlled so that the yaw rate of the vehicle matches the target yaw rate. it can.

That is, the target wheel cylinder pressure calculation processing of FIG. 5 is executed as a timer interruption processing at a predetermined cycle ΔT (for example, 5 msec), and first, in step, the steering angle detection value θ of the steering angle sensor 11 and the vehicle speed sensor 12 It reads the vehicle speed detection value V _X, then performs an operation from the transition to the vehicle speed detecting value V _X with a preset specifications of the vehicle of the above (8) to (11) below the step, the coefficient a ₁₁ ~ to calculate the a _22. Here, the constant parts _{a11v to a22v} determined by the specifications of the vehicle in the equations (8) to (11) are calculated in advance by the following equations (28) to (31).

_{a 11v = -2 (K f L} f 2 + K r L r 2) / I z ...... (28) a 12v = -2 (K f L f -K r L r) / I z ...... (29) a _21v = -2 then _{(K f L f -K r L} r) / M ...... (30) a 22v = -2 (K f + K r) / M ...... (31), the process proceeds to step, the vehicle speed detection the value V _X, advance the (18) wheel base is determined by the calculated stability factor a and specifications of the vehicle based on the equation L, and operation of the on the basis of the steering gear ratio N (17) formula By calculating the steady-state yaw rate gain H ₀ and calculating the equation (19) based on the calculated steady-state yaw rate gain H ₀ , the differential value _r (n) of the target yaw rate is calculated and further calculated. Differential value _r
From (n) and the previous value of the target yaw rate _r (n-1), the current target yaw rate _r (n) is calculated according to the following equation (32).
Is calculated and updated and stored in the target yaw rate storage area formed in the storage device 19d. _r (n) = _r (n−1) + _r (n) ΔT (32) where ΔT is a timer interrupt cycle.

Then, the process proceeds to a step, wherein the coefficients a ₂₁ and a ₂₂ calculated in the step, the coefficient b ₂ calculated in advance according to the equation (13), the target yaw rate _r (n) calculated in the step, and the lateral speed are calculated. From the previous value V _yr (n-1) of the above, the calculation of the above (21) is performed to obtain the lateral acceleration _yr (n).
Is calculated from the calculated lateral acceleration _yr (n) and the previous value of the lateral speed V _yr (n-1) according to the following formula (33) to calculate the current lateral speed V _yr (n ) Is calculated and updated and stored in the lateral speed storage area of the storage device 19d.

V _yr (n) = V _yr (n−1) + _yr (n) ΔT (33) Then, the process proceeds to step and the braking force difference ΔB _f between the left and right front wheels is calculated according to the above equation (22). Then, the calculated braking force difference ΔB _f and a proportional constant k _p calculated in advance according to equation (24)
The target differential pressure ΔP is calculated by performing the calculation of equation (25) based on the above.

Next, the process proceeds to step to determine whether or not the target differential pressure ΔP is positive. If ΔP> 0, the process proceeds to step to reduce the target cylinder pressure P ^* _FL of the front left wheel to the master cylinder pressure P ^* _FL. and sets the _MC, the timer from set to any higher value before the target cylinder pressure P ^* _FR to the master cylinder pressure P _MC and the subtraction value between the target differential pressure ΔP of the right wheel (P _MC -ΔP) or "0" After the interruption processing is completed, the process returns to the predetermined main program. When ΔP <0, the process proceeds to the step where the target cylinder pressure P ^* _FL of the front left wheel is changed to the master cylinder pressure P ^* _FL.
MC and the sum of the target differential pressure [Delta] P and sets either larger value of (P _MC + [Delta] P) or "0", set before the target cylinder pressure P ^* _FR of the right wheel to the master cylinder pressure P _MC The timer interrupt processing ends.

In the processing of FIG. 5, the processing of the step corresponds to the target yaw rate setting means, and the processing of the steps and corresponds to the target braking force calculating means.

Therefore, now assuming that continues straight ahead running state, but the vehicle speed detection value V _X from the vehicle speed sensor 12 becomes a value corresponding to the vehicle speed, with the steering angle detection value θ is zero from the steering angle sensor 11 And the previous value of the target yaw rate _r (n-
1) and the previous value V _yr (n-1) of the lateral speed are also zero. Therefore, the steady-state yaw rate gain H ₀ calculated in the step is a value corresponding to the vehicle speed, but the differential value _r (n) of the target yaw rate is determined by the steering angle detection value θ of the first term on the right side of the above equation (19). Zero and previous value of target yaw rate
_{Since r} (n-1) is also zero, it becomes zero, and therefore, the current value _r (n) of the target yaw rate also becomes zero. Accordingly, the lateral acceleration _yr (n) and the lateral velocity V _yr (n) calculated in the step also become zero, and the front wheel left-right braking force difference ΔB _f and the target differential pressure ΔP calculated in the step also become zero. , Transition from step to step. Since the vehicle is in the non-braking state, the master cylinder pressure P _MC detected by the pressure sensor 14MC is zero, target wheel cylinder pressure P ^* _FL and P ^* _FR is set to zero.

However, when the transition from the straight running state to the braking state by depressing the brake pedal 4, by the master cylinder pressure P _MC of the master cylinder 5 rises, the target wheel cylinder pressure P ^* _FL and P ^* _FR of the right and left wheels in step , it is equal to the master cylinder pressure P _MC.

On the other hand, when the vehicle is turned leftward by, for example, turning the steering wheel 10 to the left from the straight traveling state at a constant speed, the steering angle is increased from the steering angle sensor 11 in the forward direction according to the steering angle of the steering wheel 10 accordingly. Since the detected steering angle θ is output, the current value _r (n) of the differential value of the target yaw rate calculated in the step is determined based on the steady yaw rate gain H ₀ according to the vehicle speed and the detected steering angle θ. And the current value _r (n) of the target yaw rate also increases in the positive direction. Accordingly, the current value _yr (n) of the lateral acceleration calculated in the step changes in the positive or negative direction depending on the vehicle specifications and the vehicle speed, and accordingly, the current value V _yr (n) of the lateral speed. Also changes in the positive or negative direction.

Based on the above values, the braking force difference Δ
B _f and the target differential pressure ΔP are calculated. At that time, the target differential pressure ΔP
In If is negative, the process proceeds from step to step, before the target cylinder pressure P ^* _FL for the left wheel cylinder 1FL is set target differential pressure ΔP min smaller than the master cylinder pressure P _MC, front right wheel cylinder 1FR target cylinder pressure P ^* _FR is set equal to the master cylinder pressure P _MC for, by controlling the respective wheel cylinders 1FL and cylinder pressure 1FR in accordance with these, generating a proper yaw rate corresponding to the vehicle speed and steering angle can do.

When the target differential pressure ΔP is a positive value on the contrary, the process proceeds from step to step, the target cylinder pressure P ^* _FL is set to the master cylinder pressure P _MC for the previous left wheel cylinder 1FL, before the right wheel cylinder by target cylinder pressure P ^* _FR for 1FR is to be set to a value smaller target differential pressure ΔP min from the master cylinder pressure P _MC, controls the cylinder pressure of each in these depending wheel cylinders 1FL and 1FR, the vehicle speed And a yaw rate corresponding to the target yaw rate _r (n) corresponding to the steering angle and the steering angle.

Next, when the steering wheel 10 is turned right from the straight running state to the right turning state, the steering angle detection value θ of the steering angle sensor 11 becomes a negative value, so that the differential value _r (n ), The target yaw rate _r (n) becomes a negative value, but is controlled basically in the same manner as the left turn.

On the other hand, the braking force control process of FIG. 6 is individually executed on the left and right wheel sides as a timer interrupt process of a predetermined period ΔT similarly to the target cylinder pressure calculation process of FIG. FIG. 6 shows a braking force control process for the wheel cylinder 1FL on the left wheel side.

That is, it is determined in a step whether or not the brake switch 13 is in an on state. When the brake switch 14 is in an off state, it is determined that the brake is in a non-braking state, and the process proceeds to step to hold the output control signal. set to "1" to the variable T _P representing time, then set to "1" to the variable m indicating the period for monitoring an error between the actual cylinder pressure P _FL and the target cylinder pressure P ^* _FL and proceeds to step Then go to step.

In this step, it is determined whether the variable T _P is positive, “0”, or negative. If T _P > 0, the process proceeds to step and the pressure increase signal of “0” is determined. coefficient control signal CS _FL and outputs the constant current circuit 20FL, then subtracts "1" from the variable T _P and proceeds to step calculates the new coefficients T _P, to form it in the storage device 19d as After the update is stored in the storage area, the process proceeds to step, and the value obtained by subtracting "1" from the variable m is updated and stored in the variable storage area formed in the storage device 19d as a new variable m, and then the timer interrupt processing is performed. The process returns to the main program, and when the determination result of the step is T _P = 0, the process proceeds to the step and the control signal CS _FL as a holding signal of the first predetermined voltage V _S1 is obtained.
And then goes to the step. If the determination result of the step is T _P <0, the step goes to the step and the second predetermined voltage V _S2 higher than the first predetermined voltage V _S1 is output.
The outputs the control signal CS _FL as a pressure reducing signal, then migrate migrated a value obtained by adding "1" to the variable T _P and the updating stored in the variable storage area as a new variable T _P to step to step I do.

In addition, the result of the determination in the step is a brake switch.
14 when in the ON state, the vehicle is shifted to the step it is determined that a braking state, the target cylinder pressure P ^* _FL calculated by the target cylinder pressure calculating process described above is consistent with the master cylinder pressure P _MC It is determined whether or not they match, and if they match, the process proceeds to the step. If they do not match, the process proceeds to a step.

In this step, it is determined whether or not the variable m is positive. When m> 0, the process directly proceeds to the above step, and when m ≦ 0, the process proceeds to the step.

In this step, the process proceeds to the target cylinder pressure P ^* _FL and error _{^{P err (= P * FL -P}} FL) Step after calculating the current cylinder pressure detection value P _FL.

In this step calculates the variable T _P in accordance with the following equation (34) for rounding off a value obtained by dividing the error P _err in the reference value P _0.

T _P = INT (P _err / P ₀ ) (34) Next, the process proceeds to the step, sets the variable m to a predetermined value m _0, and then proceeds to the step.

Here, the processing in FIG. 6 corresponds to the braking force control means.

Therefore, in a state where the vehicle is running in the non-braking state, the brake switch 14 is in the off state.
Since the T _P> 0, the control signal CS _FL and CS _FR transition to "0" is output as a pressure signal increasing the constant current circuit 20FL and 20FR to step. For this reason, the exciting current is not output from the constant current circuits 20FL and 20FR, the electromagnetic directional control valves 3FL and 3FR maintain the normal position, and the front wheel side wheel cylinder 1FL
And 1FR are in communication with the master cylinder 5. At this time, since the brake pedal 4 is not depressed, the cylinder pressure output from the master cylinder 5 is zero, so the cylinder pressure of each of the wheel cylinders 1FL and 1FR is also zero, and a braking force is generated. The non-braking state is continued.

In this state, when the brake pedal 4 is depressed to enter the braking state, the process shifts from the step of FIG. 6 to the step, and the target cylinder pressures P ^* _FL and P ^* _FR calculated by the target cylinder pressure calculation processing of FIG. It determines whether to match the master cylinder pressure P _MC each master cylinder 5. This determination is to determine whether the vehicle is in a straight running state or a turning state. In the straight running state, as described above, the target cylinder pressures P ^* _FL and P ^* since _FR is set equal to the master cylinder pressure P _MC, the process proceeds from step to step, normal position the directional control valve 3FL and 3FR as zero both non-braking state as well as the control signal CS _FL and CS _FR described above By this, the master cylinder 5 and each wheel cylinder 1FL and 1
With FR connected, each wheel cylinder 1FL and 1FR
Increasing the cylinder pressure P _FL and P _FR to a value equal to the master cylinder pressure P _MC, it generates an equal braking force on both the wheel cylinders 1FL and 1FR.

However, when the vehicle enters the braking state in the turning state or shifts to the turning state in the braking state, in the processing of FIG. 5 described above, the target cylinder pressure P ^* _FL (or P ^* _FR ) becomes the master cylinder pressure _PMC . On the other hand, in the process for the wheel cylinder 1FL (or 1FR), the process shifts from step to step, and the variable m is set to “0” in the process of the previous step. Then, the process proceeds to the step. Therefore, each target cylinder pressure P ^* _FL (or P ^* _FR ) and the pressure sensor 14
Error P between FL (or 14FR) pressure detection value P _FL (or P _FR)
calculates _err (step), which calculates the variable T _P is divided by the set value P ₀ representing the allowable range (step), and then moves after setting the variable m to a set value m ₀ to step. At this time, the detected pressure value P _FL (or P _FR ) of each pressure sensor 14FL (or 14FR) is equal to the target cylinder pressure P ^* _FL (or P ^* _FR ).
, The variable _TP has a positive value.
The process proceeds to the step, the control signals _CSFL and _CSFR are set to zero, and the pressure increasing mode is continued. Then, in step
T _P is "1" at a time subtraction, holding it when becomes zero, the control signal CS _FL of the first predetermined voltage V _S1 shifts from step to step (or CS _FR) to the constant current circuit 20FL (or 20FR) Output as a signal. Thus, by exciting current from the constant current circuit 20FL (or 20FR) corresponding to a predetermined voltage V _S1 is output to the electromagnetic directional control valve 3FL (or 3FR), these directional control valve 3FL (or 3FR) are first 2 is switched to the switching position 2, and the connection between the wheel cylinder 1FL (or 1FR) and the master cylinder 5 is cut off.
(Or 1FR) is a holding mode in which the cylinder pressure P _FL (or P _FR ) is maintained at a constant value, and this holding mode is continued until the variable m becomes “0” in steps.

Thereafter, the variable m is "0", the process proceeds to step again, the error pressure P _err is the set pressure P variable T _P is "0" is calculated by the composed Step less than 1/2 of the _zero at this point,
The mode is shifted from step to step, and the above-mentioned holding mode is set without going through the pressure increasing mode.
1FL (or 1FR) of the cylinder pressure P _FL (or P _FR) is maintained at the target cylinder pressure P ^* _FL (or P ^* _FR).

Also, the wheel cylinder pressure P _FL (or P _FR ) of each wheel cylinder 1FL (or 1FR) is equal to the target cylinder pressure P ^* _FL (or
P ^* _FR ), the error calculated in the step
Since P _err is a negative value, the variable T _P becomes a negative value, the process proceeds from step to step outputs the control signal CS _FL (or CS _FR) of a predetermined voltage V _S2 as a vacuum signal, thus constant since the exciting current corresponding the current circuit 20FL (or 20FR) to a predetermined voltage V _S2 is supplied to the directional control valve 3FL (or 3FR), which is switched to the third switching position. Therefore, the wheel cylinder 1FL (or 1FR) is communicated with the master cylinder 5 via the hydraulic pump 7F, and the cylinder pressure P _FL (or PFR) of the wheel cylinder 1FL (or 1FR) is set.
_FR ) is in the decompression mode in which the pressure is reduced, and this mode is maintained until the variable _TP becomes “0”.

In this manner, the cylinder pressure P _FL and P _FR in the wheel cylinders 1FL and 1FR can be matched with the target cylinder pressure P ^* _FL and P ^* _FR, resulting vehicle speed and the target yaw rate in accordance with the steering angle A yaw rate that matches the optimum value can be generated. Therefore, unstable behavior due to steering in the braking state can be prevented, steering stability can be improved, and transient yaw rate characteristics can be improved.

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

In the second embodiment, the cornering power K _f , K
Focusing on the fact that _r changes due to the application of the braking force and the driving force, the target cylinder pressures P ^* _FL and P ^* _FR are calculated taking into account the change in the cornering power during braking. .

That is, the cornering force C _f and C _r of the front and rear wheels and the braking force and driving force, are related by the general concept of the friction circle, as shown in Figure 7.

Hereinafter, a method of calculating the cornering power K _r when the braking force is applied to the front wheel as an example.

First, assuming that the cornering force C _f on the front wheel side is proportional to the wheel side slip angle β, the maximum frictional force (that is, the radius of the friction circle) that can be produced by the tire is F ₀ , and the cornering force C _f is the maximum value C _fmax . if the slip angle beta beta _max, a cornering power when braking force is not applied and K _f0, relationship _{C fmax = F 0 = K f0} · β max ............ (35) is obtained.

When the braking force _Bf is applied when the equation (35) holds, the maximum cornering force C _fmax changes as in the following equation (36).

Thus, the front wheel cornering power K _f when braking force B _f is applied can be obtained by the following expression (37).

Then, when the front wheel cornering power K _f and the average value of the right and left wheels of the front wheel side, the braking force B _FL in front left, front wheel cornering power K _f when B _FR is applied is determined by the following equation (38) be able to.

Similarly, if the maximum frictional force that the rear wheel can produce is F ₀ ′ and the rear-wheel cornering power when no braking force is applied is K _r0 , the rear wheel side when the braking force _Br is applied to the rear wheel cornering power K _r can be obtained by the following equation (39).

Then, by substituting equation (23) in the first embodiment described above into equations (38) and (39), the following equation (4) is obtained.
The front wheel side cornering power K is calculated according to the equations (0) and (41).
It can be obtained _f and the rear wheel side cornering power K _r.

Then, the processing unit 19c of the control unit 16 executes the target cylinder pressure calculation processing shown in FIG. 8 in place of the target cylinder pressure calculation processing of FIG. 5 in the first embodiment described above.

That is, in steps, the vehicle speed detection value V _x , the steering angle detection value θ, and the cylinder pressures P _FL , P of the front wheel wheel cylinders 1FL, 1FR
_FR and the rear wheel side wheel cylinders 1RL, cylinder pressure 1RR P _R
Read, then calculates a front wheel cornering power K _f and the rear wheel side cornering power K _r in accordance with the equation (39) and (40) below the process proceeds to step, and then calculated in the step proceeds to step wherein the front wheel cornering power K _f and the rear wheel side cornering power K _r based on (28) - (31) by performing the calculation of the equation, the coefficient a _11v ~a _22v
, And then proceeds to the step where the same calculation as in the step of FIG. 5 is performed to obtain the coefficients a ₁₁ , a ₁₂ , a ₂₁ and a
Calculate ₂₂ . Then, step by step
The target cylinder pressures P ^* _FL and P ^* _FR are calculated by performing the same processing as described above.

According to the second embodiment, the coefficients a ₁₁ and a ₁₁ required for calculating the lateral acceleration _yr (n) and the braking force difference ΔB _f
12 , a ₂₁ , a ₂₂ , cornering power K _f included in b ₁ and b ₂
And K _r of the wheel cylinder 1FL, 1FR and 1RL, since such change in correspondence to the braking force generated by 1RR, it is possible to correct the deviation of the characteristics of the actual vehicle and a vehicle model, a more accurate The target cylinder pressures P ^* _FL and P ^* _FR can be calculated, so that the steering stability can be improved and the transient yaw rate characteristics can be improved.

Next, a modification of the second embodiment will be described with reference to FIGS. 9 and 10. FIG.

In this modified example, the radius of the tire friction circle is changed according to the wheel load movement accompanying braking.

First, the structure will be described. As shown in FIG. 10, a longitudinal acceleration sensor 21 for detecting the deceleration of the vehicle is provided in addition to the structure of the first embodiment described above. 3, except that the deceleration g is input to the input interface circuit 19a of the microcomputer 19d. The same reference numerals are given to the parts corresponding to those in FIG. Detailed description is omitted.

Next, the operation will be described. The cornering force C _f , C
_r and the braking / driving force are related by the concept of a friction circle as shown in FIG. 7 described above, and the radius of the friction circle means the maximum friction force that the tire can produce. Here, the front wheel friction circle radius of the vehicle stationary state and F _0, the front wheel friction circle radius F ₀ is a known quantity. The maximum cornering force when all the tire frictional forces are used in the lateral direction is defined as C _fmax, and the wheel cylinder pressure when all the tire frictional forces are used for the braking force is defined as _PLKF0 .

At this time, the relationship between the front wheel friction circle radius F ₀ and the maximum cornering force C _fmax is expressed by the aforementioned equation (35).

Also, the wheel cylinder pressure P _LKF0 and the front wheel friction circle radius F ₀
Can be obtained by the following equation (42) using the proportionality constant K _P of the above equation (24), if the moment of inertia of the wheel is ignored.

F ₀ = K _P・ P _LKF0 ………… (42) And the front wheel friction circle radius F ₀ , cornering power
_Considering four of K _f0 , proportional constant K _P and wheel cylinder pressure P _LKF0 as known quantities, the friction circle radius F at the time of braking, the wheel cylinder pressure P _LKF at the time of maximum braking force generation and the cornering power K _f2 at the time of braking are considered. Ask for each.

First, when a braking force is applied to the vehicle, a load shift to the front wheels occurs. The steady-state load movement amount per wheel ΔM during braking can be expressed by the following equation (43) by using the vehicle weight M, the height of the center of gravity of the vehicle h, the wheel base L, and the vehicle deceleration g.

ΔM = M · h · g / (L · 2) (43) The friction circle radius F is a ground contact load of a wheel (hereinafter referred to as a wheel load), assuming that road surface conditions are constant. ). Therefore, the friction circle radius F when the load movement of the above equation (43) occurs is calculated by the following equation (44) using the front wheel load M _f0 per wheel when the vehicle is stationary and the above load movement amount ΔM. Can be represented.

F = F ₀ (1 + ΔM / M _f0 ) (44) where M _f0 = M · L _r / (L · 2) (45) where L _r is the inner part of the wheel base. The distance from the center of gravity to the rear wheel.

Furthermore, assuming that the wheel side slip angle β _{max at} the maximum cornering force C _fmax is a constant value regardless of the wheel load, the cornering power is also proportional to the wheel load.

Thus, cornering power K _f during the load transfer
_Let K _f1 be this cornering power K _f1
6) It can be obtained by the formula.

K _f1 = K _f0 (1 + ΔM / M _f0 ) (46) The wheel cylinder pressure P _LKF when the maximum braking force is generated can be similarly obtained by the following equation (47).

P _LKF = P _LKF0 (1 + ΔM / M _f0 ) On the other hand, when the braking force B _f is applied, as shown in FIG. 7, the cornering force C _fmax that the tire can output decreases as shown in FIG. It changes like an expression.

From this equation (48), the cornering power K during braking is calculated.
_f2 can be calculated by the following equation (49).

Then, assuming that the average value of the cornering power K _f2 during braking of the left and right wheels is ▲ ▼, this average value ▲ ▼ is
The braking force B _FL and B _FR to the left and right wheels can be calculated by the following equation (50) using.

By substituting the above equations (42) and (47) into the equation (50), the average value ▲ ▼ of the cornering power during braking can be obtained by the left and right wheel cylinder pressures P _FL and P _FL.
Can be represented by the following equation (51).

Similarly, the rear wheel load per wheel when the vehicle is stationary is M _r0, and when the rear wheel load is M _r0 , the wheel cylinder pressure when the maximum braking force occurs is P _LKR0, and when the braking force is not applied if the cornering power and K _r0, cornering power K _r2 when applied is the wheel cylinder pressure P _RR the rear wheel can be calculated by the following equation (52).

Then, the processing unit 19c of the control unit 16 executes the target cylinder pressure calculation processing shown in FIG. 10 in place of the target cylinder pressure calculation processing of FIG. 8 in the second embodiment described above.

In the target cylinder pressure calculation processing of FIG.
In the processing of the figure, instead of the processing of steps and
Vehicle speed detection value V _x, the steering angle detection value theta, vehicle deceleration g, front-wheel wheel cylinder 1FL, cylinder pressure P _FL of 1FR, P _FR and the rear wheel side wheel cylinders 1RL, cylinder pressure P _R reading free steps 1RR Processing and the load movement amount ΔM according to the above equation (43).
And the processing of the step of calculating
2) performs the same processing as the FIG. 8, except that the process of calculating the front wheel cornering power K _f and the rear wheel side cornering power K _r is provided according to equation.

According to this modification, similar to the above-described second embodiment,
The cornering powers K _f and K _r included in the coefficients a ₁₁ , a ₁₂ , a ₂₁ , a _22, b ₁ and b ₂ required for calculating the lateral acceleration _yr (n) and the braking force difference ΔB _f Wheel cylinder 1FL,
1FR and 1RL, 1RR, while changing according to the braking force generated, and the friction circle radius F is changed according to the change in wheel load, the difference between the characteristics of the vehicle model and the actual vehicle more Correction can be performed accurately, and more accurate target cylinder pressures P ^* _FL and P ^* _FR can be calculated.
Steering stability can be improved, and transient yaw rate characteristics can be improved.

In each of the above embodiments, the case where the yaw rate characteristic control is performed only when the vehicle is in the braking state has been described.However, the present invention is not limited to this. For example, by using an actuator for traction control, The yaw rate characteristic control can be performed even in the non-braking state.

Further, in each of the above embodiments, the case where the braking force difference between the left and right wheels on the front wheel side is controlled has been described.
The present invention is not limited to this, and the difference between the left and right braking forces of the rear wheels or the front and rear wheels may be controlled.

Furthermore, in each of the above-described embodiments, the case where the steering angle sensor 11 is applied as the vehicle steering state detecting means has been described. However, the present invention is not limited to this. Instead of the steering angle sensor, actual steering of the wheels is performed. The angle (actual steering angle) may be detected. In this case, the aforementioned equation (3),
The steering gear ratio N in the equations (12) and (13) is omitted.

Further, in each of the above embodiments, the case where the vehicle speed sensor 12 is applied as the speed detecting means has been described. However, the present invention is not limited to this, and it is also possible to detect the wheel speed, the vehicle longitudinal acceleration and the like to calculate the vehicle longitudinal direction speed. it can.

Furthermore, in each of the above embodiments, the case where a microcomputer is applied as the braking pressure control device 16 has been described, but the present invention is not limited to this, and electronic circuits such as a comparison circuit, an arithmetic circuit, and a logic circuit are combined. It can also be configured.

〔The invention's effect〕

As described above, according to the braking force control device of the first aspect, the target yaw rate is calculated by the target yaw rate setting means based on the detected value of the steering state of the vehicle and the detected value of the longitudinal speed. The target yaw rate can be set to an optimal value according to the running state of the vehicle,
Further, in order to realize the set target yaw rate in the vehicle to be controlled, that is, in order to make the target yaw rate coincide with the yaw rate generated in the control target vehicle, the target braking force calculating means calculates the target yaw rate by the target yaw rate. Since the target braking force is calculated by a calculation based on the differential value and the steering state detection value and the vehicle model set by the vehicle specifications and the equation of motion, the vehicle traveling state such as the wheel speed and the braking force is calculated. Instead of detecting and performing feedback control, the target braking force can be calculated based on information about the vehicle yaw rate and the vehicle model, and understeer and oversteer when braking while steering is reliably suppressed. To improve steering stability and transient yaw rate characteristics. The effect is obtained that.

Further, according to the braking force control device according to claim (2),
In addition to the configuration of the braking force control device according to the first aspect, a braking pressure of each wheel is detected by a braking pressure detecting means, and a change in the front wheel side and rear wheel side cornering power is detected based on the detected braking pressure value. And the target braking force is calculated using these cornering powers, so that the deviation between the yaw rate characteristics of the vehicle model and the actual vehicle can be corrected, and the steering stability can be further improved. As a result, the effect that the transient yaw rate characteristic can be further improved is obtained.

Furthermore, according to the braking force control device according to claim (3), in the configuration of the braking force control device according to claim (2), the target braking force calculation means uses the braking pressure detection value and the load shift during braking. The change of the front wheel side and rear wheel side cornering power is estimated based on the change of the friction circle radius accompanying the above, so that the difference between the vehicle model and the actual yaw rate characteristic of the actual vehicle can be more accurately corrected. Is obtained.

[Brief description of the drawings]

1 (a) and 1 (b) are schematic diagrams showing a basic configuration of the present invention, FIG. 2 is a system diagram showing a first embodiment of the present invention, and FIG. 3 shows an example of a braking pressure control device. Block Diagram,
FIG. 4 is an explanatory view of a vehicle motion model, FIGS. 5 and 6 are flowcharts each showing an example of a processing procedure of a braking pressure control device, and FIG. 7 is a frictional chart for explaining a second embodiment of the present invention. FIG. 8 is an explanatory diagram showing the concept of a circle, FIG. 8 is a flowchart showing a processing procedure corresponding to FIG. 5 of the braking pressure control device according to the second embodiment of the present invention, and FIG. 9 shows a modification of the second embodiment. FIG. 3 is a block diagram of a braking pressure control device corresponding to FIG. 3;
FIG. 10 is a flowchart showing a processing procedure corresponding to FIG. 8 of the braking pressure control device in the modification. In the figure, 1FL-1RR is a wheel cylinder (braking means), 2 is an actuator, 3FL-3R is an electromagnetic direction switching valve, 4 is a brake pedal, 5 is a master cylinder, 11 is a steering angle sensor (steering state detecting means), 12 Is a vehicle speed sensor (speed detecting means), 13 is a brake switch, 14FL-14MC are pressure sensors (braking pressure detecting means), 16 is a braking pressure control device, 19 is a microcomputer, and 21 is a longitudinal acceleration sensor.

Claims (3)

(57) [Claims] 1. A steering state detecting means for detecting a steering state of a vehicle, a speed detecting means for detecting a longitudinal speed of the vehicle,
The target yaw rate setting means for setting the target yaw rate of the vehicle based on the detection values of the steering state detecting means and the speed detecting means, and the braking forces of the left and right braking means disposed on at least one of the front wheels and the rear wheels are independently controlled. Controllable braking force control means, and at least one of the left and right target braking forces of the front wheels and the rear wheels required to achieve the target yaw rate set by the target yaw rate setting means in the vehicle to be controlled. A target yaw rate, a differential value of the target yaw rate, a detection value of the steering state detection unit, and a target braking force calculation unit that calculates by a calculation based on a vehicle model set in advance by vehicle specifications and a motion equation, The braking force control means controls the braking force of the braking means so as to match the target braking force calculated by the target braking force calculation means. Brake force control apparatus which is characterized in that. 2. A steering state detecting means for detecting a steering state of a vehicle, a speed detecting means for detecting a longitudinal speed of the vehicle,
Braking pressure detecting means for detecting the braking pressure of the four-wheel braking means;
The target yaw rate setting means for setting the target yaw rate of the vehicle based on the detection values of the steering state detecting means and the speed detecting means, and the braking forces of the left and right braking means disposed on at least one of the front wheels and the rear wheels are independently controlled. A controllable braking force control means, and a target braking force on at least one of the left and right front wheels and the rear wheels necessary for realizing the target yaw rate set by the target yaw rate setting means on the vehicle to be controlled are determined in advance in the vehicle. And a target braking force calculating means for calculating by a calculation based on a vehicle model set based on a source and an equation of motion, wherein the target braking force calculating means calculates at least a value corresponding to a cornering power of a wheel in the vehicle model by the braking. The braking force is changed according to the detection value of the pressure detecting means, and the braking force control means controls the braking force of the braking means by the target braking force calculating means. Braking force control device and controls to match the calculated target braking force. 3. The target braking force calculating means is configured to change a value corresponding to a cornering power based on a value detected by a braking pressure detecting means and a change in a radius of a friction circle accompanying a load movement. Item (2): The braking force control device.