JPH0524422A - Braking force controller - Google Patents

Abstract

PURPOSE:To improve transient characteristics and restrain insufficient braking force by computing the stationary gain of a target yaw rate according to a change in cornering power with load change at the time of braking. CONSTITUTION:A target yaw rate psi'r is computed from the steering condition detection value theta and the front/rear direction speed detection value Vx of a vehicle-(S3) and accurate cornering power is computed from a braking pressure detection value for each wheel and a wheel load detection value-(S5). Target braking forces P*FL, P*FR are then obtained by the calculation based on a vehicle model using the above cornering power-(S12, S13) so that the target yaw rate psi'r and a yaw rate actually generated at a vehicle may meet each other, according to which the braking force of the wheel is controlled. Moreover, the target yaw rate psi'r is compensated with cornering power-(S7).

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a braking force control device capable of improving the steering stability of a vehicle during braking.

[0002]

2. Description of the Related Art As a conventional braking force control device, for example, as disclosed in Japanese Utility Model Laid-Open No. 59-155264, there is one that controls the vehicle yaw characteristic by the left and right brake differential pressure. Specifically, when the driver's steering angle is equal to or greater than a predetermined value, the control is performed so as to delay the pressure increase timing of the turning outer wheel to improve the turning ability during braking.

However, the above-mentioned conventional braking force control device does not consider that the yaw rate generated by the front wheel steering and the left-right braking force difference depends on the vehicle speed, and it is difficult to control the yaw rate to an appropriate value. However, there is an unsolved problem that it is difficult to improve the transient characteristics of the generated yaw rate. In order to solve such a problem, the present applicant has previously proposed the braking force control device described in Japanese Patent Application No. 2-219367. According to this braking force control device, the target yaw rate is set from the vehicle speed and the steering angle, and the target left-right braking force difference required to match the target yaw rate with the yaw rate generated in the actual vehicle is set in advance in the vehicle specifications and It is calculated by calculation based on the vehicle model set by the equation of motion, and the left and right braking forces calculated from this target left and right braking force difference are controlled to match the actual left and right braking force. There is an advantage that the transient characteristic of the generated yaw rate is improved. In this braking force control device, the target braking force changes the cornering power of the wheels in the vehicle model based on the braking pressure and the friction circle radius during braking, and the target braking force is changed using the cornering power. Is calculated

[0004]

However, in the conventional braking force control device, the predetermined yaw characteristic determined by the steering angle and the vehicle speed is obtained regardless of whether the steady-state gain of the target yaw rate is applied when the vehicle is braking or not braking. However, when turning while braking, the cornering power of the wheels changes due to the movement of the load to the front wheels due to braking during turning, and the steady yaw rate gain of the vehicle in the uncontrolled state is oversteered. Therefore, a deviation always occurs with respect to the fixed target value during braking, and control is always performed so that the deviation becomes zero.

Therefore, the frequency of operation of the system, especially the control valve, may be increased, which may adversely affect the durability, and the system becomes expensive in consideration of this.
Further, in a normal vehicle, when turning while braking, the steer characteristic changes in the oversteer direction as described above, so for a driver who is accustomed to this or a driver who intends to actively use this characteristic, It is not preferable to control to the same steer characteristics as when not braking even when braking.

Further, as a method for obtaining the braking force difference between the left and right wheels, the braking fluid pressure on one wheel is reduced below the master cylinder pressure,
When adopting the method of increasing the braking fluid pressure of the opposite wheel from the master cylinder pressure, a special pressure increasing means is required, which leads to system complexity and cost increase. The method of decompressing only the side is advantageous without the above problem, but in this case, there is a problem that deceleration is insufficient because the total braking force of the left and right wheels is reduced.

The present invention has been made by paying attention to the above-mentioned problems. The steady gain of the target yaw rate is changed according to the change of the cornering power accompanying the load change at the time of braking, which is inexpensive and meets the driver's request. It is an object of the present invention to provide a braking force control device that can achieve high steering stability while suppressing the deceleration insufficiency while providing the characteristics that meet the requirements.

[0008]

In order to achieve the above object, the braking force control apparatus of the present invention has a steering state detecting means for detecting the steering state of a vehicle and a vehicle, as shown in the basic configuration of FIG. Of the front-rear direction speed, a target yaw rate setting means for setting a steady gain of the target yaw rate of the vehicle by inputting signals from the steering state detection means and the speed detection means, and the target yaw rate setting means. Calculate the target braking force on the left and right of at least one of the front and rear wheels necessary to realize the set target yaw rate on the vehicle to be controlled by calculation based on the vehicle model set in advance by the vehicle specifications and the equation of motion Target braking force calculation means and independent braking forces of the left and right braking means disposed on at least one of the front wheels and the rear wheels so as to match the target braking force. In a braking force control device provided with a possible braking force control means, a wheel load detection means for detecting a load of each wheel is provided, and the target yaw rate setting means and the target braking force calculation means include at least the wheel load. The steady gain of the target yaw rate is set and the target braking force is calculated according to the cornering power of the wheel estimated from the detection value of the detection means.

[0009]

In the braking force control device of the present invention, the target yaw rate setting means calculates the target yaw rate ψ'r based on the steering state detection value of the vehicle, for example, the steering angle detection value, and the longitudinal speed of the vehicle, for example, the vehicle speed. Then, in order to match the target yaw rate ψ′r with the yaw rate actually generated in the vehicle, the target braking force calculation means performs a calculation based on the vehicle model set by the vehicle specifications and the equation of motion, and A target braking force that causes the braking force control means to generate a braking force difference is calculated. At this time, a value corresponding to the cornering power in the vehicle model is calculated based on the detected value of the load on each wheel, and the steady gain of the target yaw rate and the target braking force are calculated using this cornering power. 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 braking force of the vehicle can be controlled to a more appropriate value. The steering stability is improved and the transient braking characteristics are improved. Furthermore, since the steady-state gain of the target yaw rate approaches the value that actually occurs in the vehicle, the control input and system operation frequency in the steady state are reduced.

[0010]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 2 is a hydraulic / electrical system diagram showing a first embodiment of the present invention. In the figure, 1FL and 1FR are wheel cylinders 1RL and 1R as braking means attached to the front wheels.
R is a wheel cylinder as a braking means attached to the rear wheels, and these wheel cylinders 1FL to 1RR
The brake hydraulic pressure supplied to the actuator is controlled by the actuator 2.

This actuator 2 is a three-port three-position electromagnetic directional control valve 3FL and 3FR for individually controlling the cylinder pressures of the front wheel side wheel cylinders 1FL and 1RR.
And a 3-port 3-position electromagnetic directional control valve 3R for simultaneously controlling the wheel cylinders 1RL and 1RR on the rear wheel side. Then, P of the electromagnetic directional control valves 3FL and 3FR
The port is connected to one system of the two-system master cylinder 5 connected to the brake pedal 4, the A ports of the electromagnetic directional control valves 3FL and 3FR are individually connected to the wheel cylinders 1FL and 1FR, and the B port is electrically operated. It is connected to one system of the master cylinder 5 via a hydraulic pump 7F which is rotationally driven by a 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 of the electromagnetic directional control valve 3R is connected to the other of the master cylinder 5 via a hydraulic pump 7R which is rotationally driven by an electric motor (not shown). Connected to the grid. Further, the electromagnetic directional control valves 3FL and 3FL
An accumulator 8F is connected to a pipe between the P port of the FR and the hydraulic pump 7F, a reservoir tank 9F is connected to a pipe between the B port and the hydraulic pump 7F, and the P of the electromagnetic directional control valve 3R is similarly connected. An accumulator 8R is connected to a pipe line between the port and the hydraulic pump 7R, and a reservoir tank 9R is connected to a pipe line between the B port and the hydraulic pump 7R.

In each of the electromagnetic directional control valves 3FL to 3R, the brake fluid of the wheel cylinders 1FL to 1RR is directly connected to the master cylinder 5 and the wheel cylinders 1FL to 1RR at the normal first switching position. The pressure is set to a value corresponding to the master cylinder 5, and the wheel cylinders 1FL to 1RR are shut off from the master cylinder 5 and the hydraulic pumps 7F and 7R at the second switching position. By setting the holding state for holding the brake fluid pressure and further connecting the wheel cylinders 1FL to 1RR and the master cylinder 5 via the hydraulic pump at the third switching position, the hydraulic oil in the wheel cylinders 1FL to 1RR is released. The pressure reducing state is returned to the master cylinder 5 side, and the switching positions of these are changed to the braking pressure control device 1 described later. It is switching control by the current value of the three stages supplied by.

On the other hand, the vehicle has a steering wheel 1
When a steering angle of 0 is detected, a zero voltage is detected when the steering wheel 10 is in the neutral position, a negative voltage corresponding to the steering angle when the steering wheel 10 is turned right from the neutral position, and a left voltage when the steering wheel 10 is turned left from the neutral position. A steering angle sensor 11 as a steering state detecting means for outputting a steering angle detection value θ which becomes a positive voltage according to the steering angle is provided, and a speed detection for outputting a vehicle speed detection value V _X according to the vehicle speed. Vehicle speed sensor 1 as a means
2 is attached, and a brake switch 13 for detecting the depression state of the brake pedal 4 is attached, and each wheel cylinder 1FL, 1FR, 1RL, 1R
Pressure sensor 1 that outputs pressure detection values P _FL , P _FR , P _R, and P _MC according to the cylinder pressure of R and the master cylinder 5.
4FL, 14FR, 14R and 14MC are attached,
Further, wheel load sensors 15F and 15R configured by, for example, load cells, which output load detection values W _f and W _r corresponding to the wheel loads of the front wheels and the rear wheels, respectively, are mounted, and the detection values of these sensors are used for braking pressure control. Input to the device 16.

As shown in FIG. 3, the braking pressure control device 16 includes sensors 11, 12, 13, 14FL to 14MC,
, 15F and 15R, and the control signals CS _FL , CS _FR and CS _R output from the microcomputer 19 are individually input to the electromagnetic directional control valves 3FL and 3FR. And floating constant current circuits 20FL, 20FR and 20R for driving the 3R solenoids.

The microcomputer 19 has an input interface circuit 19 having at least an A / D conversion function.
a, output interface circuit 1 having D / A conversion function
9b, an arithmetic processing unit 19c and a storage unit 19d,
The arithmetic processing unit 19c executes the processing of FIG. 5 based on the steering angle detection value θ from the steering angle sensor 11, the vehicle speed detection value V _X from the vehicle speed sensor 12 and the master cylinder pressure detection value P _MC from the pressure sensor 14MC. The target wheel cylinder pressures P ^* _FR and P ^* _FL are calculated as the target braking force for the left and right front wheels, and the target wheel cylinder pressures P ^* _FR and P ^* _FL and the cylinder pressure detection values of the pressure sensors 14FR, 14FL and 14MC are calculated. Figure 6 based on P _FR , P _FL and P _MC
And the electromagnetic directional control valve 3 of the actuator 2 is executed.
The control signals CS _FL and CS _FR for controlling the FL and 3FR are output, and the control signal CS _{R of} zero is always output to the electromagnetic directional control valve 3R.

Next, the operation of the above embodiment will be described. First, the control principle of this embodiment will be described. As shown in FIG.
When considering the degree of freedom, the equation of motion can be expressed by the following equations 1 and 2. I _Z · φ ′ (t) = C _f · L _f −C _r · L _r + T _f · (B _FL (t) −B _FR (t)) / 2 ……… (1) MV · Vy ( _{t) = 2 (C f +} C r) -M · Vx (t) · ψ '(t) ......... (2) where I _Z is a vehicle yaw inertia moment, ψ' (t) 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) is the left front wheel braking force, B _FR (t) is the right front wheel braking force, M
Is the vehicle weight, Vy (t) is the vehicle lateral velocity, V'y (t) is the vehicle lateral acceleration, and Vx (t) is the vehicle longitudinal velocity.
Further, C _f and C _r are cornering forces of the front wheels and the rear wheels, and can be expressed by the following formulas 3 and 4.

C _f = K _f {θ (t) / N− (Vy (t) + L _f · ψ (t)) / Vx (t)} (3) C _r = −K _r (Vy (t ) −L _r ψ ′ (t)) / Vx (t) (4) where θ (t) is the steering angle, N is the steering gear ratio, and K _f
Is the front wheel cornering power, and K _r is the rear wheel cornering power. Substituting these equations 3 and 4 into the equations 1 and 2, and considering them as differential equations regarding the yaw rate ψ '(t) and the lateral velocity Vy (t), they should be expressed by the following equations 5 and 6. You can

Ψ ″ (t) = a ₁₁ / Vx (t) · ψ ′ (t) + a ₁₂ / Vx (t) · Vy (t) + b ₁ · θ (t) + b _pl · ΔB _f (t) ... …… (5) V'y (t) = (a ₂₁ / Vx (t) -Vx (t)) ・ ψ '(t) + a ₂₂・ Vx (t) ・ Vy (t) + b ₂・ θ (t ) ……… (6) However, ΔB _f (t) = B _FL (t) −B _FR (t) …… (7.1) a ₁₁ = −2 (K _f · L _f · L _f + K _r · L _r・ L _r ) / I _Z ...... (7.2) a ₁₂ = -2 (K _f · L _f −K _r ⋅ L _r ) / I _Z …… (7.3) a ₂₁ = -2 (K _f · L _f − K _r · L _r ) / M …… (7.4) a ₂₂ = −2 (K _f + K _r ) / M …… (7.5) b ₁ = 2 · K _{f Ũ} L _f / (I _Z Ũ) …… (7.6) b ₂ = 2 · K _f / (M · N) …… (7.7) b _pl = T _f / (2 · I _z ) …… (7.8) Considering a normal vehicle, front wheel braking force difference ΔB _f Since (t) is zero, the steering angle θ is neglected by ignoring the term of ΔB _f (t) in the above equation (5).
Consider the transfer function of the yaw rate ψ '(t) with respect to (t).

[0020] First, Equation 5 than Vy (t) = (Vx ( t) · ψ "(t) -a 11 · ψ '(t) -b 1 · Vx (t) · θ (t)) / a 12 ... ... (8) by differentiating the formula 8 of both sides, it expressed using a differential operator S, SVy = (SVx ψ " -a 11 · Sψ '-b 1 · SVx θ) / a 12 ......... ( 9) further SVy by transforming the formula (9) = {(V'xSψ '+ Vx S 2 ψ') -a 11 · Sψ '-b 1 (V'xθ + Vx Sθ)} / a 12 ......... (10) said 6 From the equations 9, 9 and 10, the input / output transfer function corresponding to the change of Vx is expressed by the following equation 11.

[0021] The transfer function of Equation 15 is in the form of (first-order) / (second-order)
And the vehicle longitudinal velocity V _XBecomes larger, the steering angle input θ
The generated yaw rate ψ '(t) for (S) becomes oscillatory and
It can be seen that both maneuverability and stability deteriorate.

Therefore, for example, if the target yaw rate ψ'r (S) is a first-order lag system with no overshoot or undershoot with respect to the steering angle input θ (S), and the steady value is set equal to that of a normal vehicle. , The target yaw rate ψ'r (S) is the following 1
It can be expressed by two expressions. ψ'r (S) = H ₀ · θ (S) / (1 + τ S) ... (12) However, H ₀ is a steady yaw rate gain, and the transfer function 11
By defining the differential term of the equation as 0, it is defined by the following 13 equations.

[0023] Next, the calculation of the cornering powers K _f and K _r expected in an actual vehicle will be described. Generally, the cornering forces C _f and C _r of the front wheels and the rear wheels are related to the braking force and the driving force by the concept of a friction circle as shown in FIG. Hereinafter, a method of calculating the cornering power K _f when the braking force is applied will be described taking the front wheels as an example.

First, the cornering force C _f on the front wheel side
Is assumed to be proportional to the wheel side slip angle β, the maximum frictional force (that is, the radius of the friction circle) that the tire can produce is F ₀ , and the side slip angle β when the cornering force C _f has the maximum value C _fmax is β _max and the braking force is _Letting K _{f0 be} the cornering power when is not added, the relationship of C _fmax = F ₀ = K _f0 · β _max (15) can be obtained.

When the braking force B _f is applied when the expression 15 is satisfied, the maximum cornering force C _fmax is 16
It changes like the formula. ^{_{C fmax = (F 0 2 -B}} f 2) 1/2 = F 0 (1- (B f / F 0) 2) 1/2 = {K f0 (1- (B f / F 0) 2) 1 ^{/ 2} } β _max (16) Therefore, the cornering power K _f when the braking force B _f is applied can be obtained by the following 17 equation.

K _f = K _f0 (1- (B _f / F ₀ ) ² ) ^1/2 (17) Then, if the front wheel side cornering power K _f is an average value of the front left and right wheels, The front wheel side cornering power K _f when the braking forces B _FL and B _FR are applied to the left and right can be obtained by the following 18 equation. _{K f = K f0 {(1-} (B FL / F 0) 2) 1/2 + (1- (B FR / F 0) 2) 1/2} / 2 ......... (18) Similarly, after the maximum frictional force can be generated by the wheels F ₀ ', the wheel side cornering power after when the braking force is not applied if K _r0, wheel cornering power K _r after when braking force B _R is applied to the rear wheel Can be calculated by the following equation (19).

[0027] _{K r = K r0 (1- (} B R / F 0 ') 2) 1/2 ......... (19) Further, cornering power varies with the load applied to the tire. At this time, from the tire characteristics, the sideslip angle β
If is a sufficiently small region, the cornering power is proportional to the wheel load. Therefore, with respect to the front wheel side cornering power K _fo 'and the rear wheel side cornering power K _r0 ' at a static load, the front wheel side cornering power K _f 'and the rear wheel side in consideration of load changes according to the following formulas 20 and 21. The cornering power K _r 'can be obtained.

K _f '= K _fo ' -W _f / W _fo ... (20) K _r '= K _r0 ' -W _r / W _r0 ... (21) where W _f is the front wheel load, W _r is the load on the rear wheels, W _fo is the load on the front wheels when stationary, and W _r0 is the load on the rear wheels when stationary. The front wheel side cornering power K _f and the rear wheel side cornering power K _r in consideration of braking and load movement can be obtained from the above equations 20 and 21 by the following equations 22 and 23.

K _f = K _f0 · W _f / W _fo · {(1- (B _FL / F ₀ ) ² ) ^1/2 + (1- (B _FR / F ₀ ) ² ) ^1/2 } / 2 ……… (22) K _r = K _r0・ W _r / W _r0・ (1- (B _RR / F ₀ ') ² ) ^1/2 ……… (23) The front wheels obtained from these formulas 22 and 23. The side cornering power K _f and the rear wheel side cornering power K _r are described in the above 7.
2 to 7.6, and by using them to set the steady-state yaw rate gain H ₀ shown in the above equation 13 and the target yaw rate ψ′r (t) shown in the above equation 12, the target yaw rate during braking or during load movement. ψ'r (t) is corrected at any time.

Next, using the braking force difference ΔB _f (t) between the left and right front wheels, the yaw rate ψ ′ (t) generated by the vehicle is changed to the target yaw rate ψ′r.
How to match (t) is explained. The differential value ψ ″ r (t) of the target yaw rate can be obtained by the following formula 24, which is a modification of formula 12 above: ψ ″ r (t) = H ₀ · θ (t) / τ−ψ′r (t) / Τ ……… (24) Assuming that the yaw rate ψ '(t) generated by the steering angle input θ (t) and the front wheel left-right braking force difference ΔB _f (t) matches the target yaw rate ψ'r (t). , Each differential value ψ "(t), ψ" r (t)
Also matches. Therefore, ψ "r (t) = ψ" (t), ψ'r (t) = ψ '
(t) and the lateral velocity Vy (t) when the above assumption is satisfied are defined as Vyr (t), and by substituting these into the above equations 5 and 6, the following 25 equations and Equation 26 can be obtained.

Ψ ″ r (t) = a ₁₁ / Vx (t) · ψ′r (t) + a ₁₂ / Vx (t) · Vyr (t) + b ₁ · θ (t) + b _pl · ΔB _f (t ) ……… (25) Vyr ′ (t) = (a ₂₁ / Vx (t) −Vx (t)) · ψ′r (t) + a ₂₂ · Vx (t) · Vyr (t) + b ₂ · θ (t) ... (26) Then, by substituting the above equation 24 into the above equation 25, the braking force difference ΔB _f (t) between the left and right front wheels can be obtained by the following equation 27.

ΔB _f (t) = (ψ ″ r (t) −a ₁₁ / Vx (t) · ψ′r (t) −a ₁₂ / Vx (t) · Vyr (t) −b ₁ · θ ( t)) / b _pl ………… (27) In order to generate the braking force difference ΔB _f (t) between the left and right front wheels, which is obtained by the equation (27), it is necessary to generate a differential pressure between the wheel cylinder pressures on the left and right front wheels. Well, the relationship between the wheel cylinder pressure P and the braking force B _f is, if the moment of inertia of the wheel is ignored,
It can be calculated by the following formula 28.

[0033] _{B f = 2 · μ p ·} A p · r p · P / R = k p · P ......... (28) k p = 2 · μ p · A p · r p / R ......... ( 29) where k _p is a proportional constant between the wheel cylinder pressure and the braking force, μ _p is the friction coefficient between the brake pad and the disc rotor, A _p is the wheel cylinder area, r _p is the disc rotor effective radius, and R is the tire. Is the radius.

Therefore, if the target differential pressure of the front and left wheel cylinder pressures is ΔP (t), this target differential pressure ΔP
(t) can be represented by ΔP (t) = ΔB _f (t) / k _p (30). Then, from the target differential pressure ΔP (t) and the master cylinder pressure P _MC (t) obtained by the above formula 30, the target wheel cylinder pressures P ^* _FL (t) and P ^* _FR (t) on the left and right of the front wheels are calculated as follows. It is calculated according to the equations 31 and 32.

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)) ......... (31) 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 )) (32) Therefore, the arithmetic processing unit 19c of the microcomputer 19 executes the target wheel cylinder pressure arithmetic processing of FIG. 6 and the braking force control processing of FIG. 7 to control the left and right front wheels. Power can be controlled to match the yaw rate of the vehicle to the target yaw rate.

That is, the target wheel cylinder pressure calculation of FIG.
The processing is a timer interrupt for each predetermined cycle ΔT (for example, 5 msec)
It is executed as a process, and first, in step S1, the steering angle sensor is
Steering angle detection value θ of the vehicle 11 and vehicle speed detection of the vehicle speed sensor 12
Value V_XRead, then move to step S2
Cylinder pressure sensors 14FL, 14FR, 14R and
And master cylinder pressure sensor 14MC cylinder pressure
Detection value P_FL, P_FR, P _RAnd P_MCRead, then S3
Is calculated based on the vehicle speed detection value Vx and the above equation 14 in advance.
Determined according to the stability factor A and the specifications of the vehicle
For the specified wheel base L and steering gear ratio N
Based on the above equation 13, the steady yaw rate gay is calculated.
H₀And the calculated steady-state yaw rate
In H₀By performing the calculation of Equation 12 based on
Calculate the target yaw rate ψ'r (t). Here, the above 7.2
Equation-Determined by the vehicle specifications in Equation 7.6
Number a₁₁~ A_{twenty two}Is calculated in advance.

Next, in step S4, the wheel load detection values W _f and W _r of the front and rear wheels output from the wheel load sensors 15F and 15R are read, and in step S5, the wheel load detection values W _f and W _r and preset maximum frictional forces F ₀ and F ₀ ′ of front and rear wheels, front and rear wheel load values W _f0 and W _r0 at rest
The above equations (22) and (23) are performed using the above equations to calculate the cornering powers K _f and K _r during braking or load movement.

In the following step S6, during the braking or
Cornering power K when moving loads _f, K_rUsing
Note: Determined according to the specifications of the vehicle in Formulas 7.2 to 7.6
Coefficient a₁₁~ A_{twenty two}Is corrected, and then to step S7
Transition to these cornering powers K_f, K_rUsing
And the steady-state yaw rate gain H of the above equation 13₀Correction of
U

Then, the process proceeds to step S8, where the above 24 is calculated based on the corrected steady-state yaw rate gain H _0.
The differential value ψ "r (n) of the target yaw rate is calculated by performing the equation calculation, and the calculated differential value ψ" r (n) and the previous value ψ'r (n-1) of the target yaw rate are calculated. Then, the current target yaw rate ψ′r (n) is calculated according to the following equation 33, and this is updated and stored in the target yaw rate storage area formed in the storage device 19d.

Ψ'r (n) = ψ'r (n-1) + ψ "r (n) ∆T ... (33) where ∆T is the timer interrupt period.
Shift to step S9, and the coefficient a corrected in step S6
_{twenty one}, A _{twenty two}, B₂And the target value calculated in step S8.
-The rate ψ'r (n) and the previous value of lateral velocity Vyr (n-1)
The above 26 calculation is performed to calculate the lateral acceleration Vyr '(n).
Then, the calculated lateral acceleration Vyr '(n) and lateral velocity
From the previous value Vyr (n-1) of
The present lateral velocity Vyr (n) is calculated and stored in the storage device 19
It is updated and stored in the lateral velocity storage area d.

Vyr (n) = Vyr (n-1) + Vyr '(n) .ΔT ... (34) Then, the process proceeds to step S10, and the braking force difference ΔB _f between the left and right front wheels is calculated according to the above equation 27. Then, based on the calculated braking force difference ΔB _f and the proportional constant k _p calculated in advance according to the 29 equation, the equation 30 is calculated,
The target differential pressure ΔP is calculated.

Then, the process proceeds to step S11, where the target
It is determined whether or not the differential pressure ΔP is positive, and if ΔP> 0,
If so, move to step S12 and set the target front wheel
Pressure P^* _FLIs the master cylinder pressure P_MCWhen set to
To the target cylinder pressure P of the front right wheel ^* _FRMaster cylinder pressure
P_MCAnd the target differential pressure ΔP (P_MC-ΔP) or
Timer interrupt processing after setting to the larger value of "0"
Ends and returns to the specified main program, and ΔP <0
If it is, the process proceeds to step S13 and the front left wheel
Target cylinder pressure P^* _FLIs the master cylinder pressure P_MCAnd the target difference
Value added to pressure ΔP (P _MC+ ΔP) or “0”, whichever is greater
The target cylinder pressure P of the front right wheel is set with the threshold value.^*
_FRIs the master cylinder pressure P_MCAfter setting to
End the reason.

In the process of FIG. 6, steps S3 to S3.
The processing of 5, 7, and 8 corresponds to the target yaw rate setting means,
The processing of steps S6 and 9 to 13 corresponds to the target braking force calculation means. Therefore, if it is assumed that the straight traveling state is continued, the vehicle speed detection value Vx from the vehicle speed sensor 12 is obtained.
Is a value corresponding to the vehicle speed, but the steering angle detection value θ from the steering angle sensor 11 is zero, and the previous value ψ′r (n-1) of the target yaw rate and the previous value Vyr (n of the lateral speed are -1) is also zero. Therefore, the steady-state yaw rate gain H ₀ corrected in step S7 becomes a value according to the vehicle speed, but the differential value ψ ″ r (n) of the target yaw rate is the first value on the right side of the equation 24.
Since the steering angle detection value θ of the term is zero and the previous value ψ′r (n−1) of the target yaw rate is also zero, the current value ψ′r (n) of the target yaw rate is also zero. Accordingly, the lateral acceleration Vyr (n) and the lateral velocity Vyr (n) calculated in step S9 also become zero, and the front wheel left-right braking force difference ΔB _f and the target pressure difference ΔP calculated in step S10 also become zero. Therefore, the process proceeds from step S11 to step S13. Here, since the vehicle is in the non-braking state, the master cylinder pressure P _MC detected by the pressure sensor 14MC is zero, so the target wheel cylinder pressures P ^* _FL and P ^* _FR are set to zero.

However, when the brake pedal 4 is depressed from the straight running state to the braking state, the master cylinder pressure P _MC of the master cylinder 5 rises.
In step S13, the target wheel cylinder pressure P ^* for the left and right wheels
_FL and P ^* _FR are set equal to the master cylinder pressure P _MC . On the other hand, when the vehicle turns left by turning the steering wheel 10 to the left from the straight-ahead constant-speed running state, the steering angle sensor 11 increases accordingly in the positive direction according to the steering angle of the steering wheel 10. The steering angle detection value θ is output. Further, in step S7, the steady-state yaw rate gain H ₀ is corrected to a value according to the wheel load movement that occurs in the left turning state. Therefore, the current value ψ ″ r (n) of the differential value of the target yaw rate calculated in step S8 becomes a value corresponding to the steady yaw rate gain H ₀ and the steering angle detection value θ according to the vehicle speed and the left turning state, The current value ψ′r (t) of the target yaw rate also increases in the positive direction, and the current value Vyr ′ (n) of the lateral acceleration calculated in step S9 is positive depending on the vehicle specifications and the vehicle speed. The current value Vyr (n) of the lateral speed also changes in the positive direction or the negative direction.

Based on the above values, in step S10, the front wheels
Left-right braking force difference ΔB_fAnd the target differential pressure ΔP is calculated.
At this time, if the target differential pressure ΔP is a negative value, step S1
1 to step S13, the front left wheel cylinder
Target cylinder pressure P for da 1FL^* _FLIs Master Siri
Pressure P_MCThe target differential pressure ΔP is set smaller, and the front right
Target cylinder pressure P for oil cylinder 1FR^* _FRBut
Master cylinder pressure P _MCIs set equal to
Cylinder pressure of each wheel cylinder 1FL and 1FR
Control the vehicle speed, steering angle and load movement.
The proper yaw rate can be generated.

On the contrary, when the target differential pressure ΔP is a positive value, the process proceeds from step S11 to step S12, and the target cylinder pressure P ^* _FL for the front left wheel cylinder 1FL is set to the master cylinder pressure P _MC. , The target cylinder pressure P ^* _FR for the front right wheel cylinder 1FR is set to a value smaller than the master cylinder pressure P _MC by the target differential pressure ΔP, and the cylinder pressures of the wheel cylinders 1FL and 1FR are set accordingly. By controlling, it is possible to generate a yaw rate that matches the target yaw rate ψ′r (n) according to the vehicle speed, the steering angle, and the load movement.

Next, when the steering wheel 10 is turned to the right by turning the steering wheel 10 from the straight running state, the steering angle detection value θ of the steering angle sensor 11 becomes a negative value, so that the differential value of the target yaw rate is obtained. ψ ″ r (n) and the target yaw rate ψ′r (n) have negative values, but the control is basically performed in the same manner as the left turn. On the other hand, the braking force control process of FIG. Similar to the target cylinder pressure calculation process, the left and right wheels are individually executed as a timer interrupt process of a predetermined cycle ΔT Note that Fig. 7 shows only the braking force control process for the left wheel side wheel cylinder 1FL.

That is, in step S14, it is determined whether or not the brake switch 13 is in the ON state. When the brake switch 13 is in the OFF state, it is determined that the brake is not applied, the process proceeds to step S15, and the output A variable T _P representing the holding time of the control signal to be set is set to "1", and then the process proceeds to step S16 to make a variable m representing a cycle for monitoring an error between the target cylinder pressure P ^* _FL and the actual cylinder pressure P _FL. Is set to "1" and the process proceeds to step S17.

In step S17, it is determined whether the variable T _P is positive, “0”, or negative. If T _P > 0, the process proceeds to step S18 and “0” is set. The control signal CS _FL as the "pressure increasing signal" is output to the constant current circuit 20FL, then the process proceeds to step S19, "1" is subtracted from the variable T _P to calculate a new coefficient T _P , and this is stored. After updating and storing in the coefficient storage area formed in the device 19d, the process proceeds to step S20, and the value obtained by subtracting “1” from the variable m is set as a new variable m in the storage device 19
After updating and storing in the variable storage area formed in d, the timer interrupt processing is ended and the process returns to the main program. When the determination result in step S17 is T _P = 0, the process proceeds to step S21 and the first After the control signal CS _FL as the holding signal of the predetermined voltage V _S1 is output, the process proceeds to step S20, and when the determination result of step S17 is T _P <0, the process proceeds to step S22 and the first predetermined value is obtained. outputs the control signal CS _FL as a pressure reducing signal higher than the voltage V _S1 second predetermined voltage V _S2, then a value obtained by adding "1" to the variable T _P and proceeds to step S23 as a new variable T _P After updating and storing in the variable storage area, the above step S2
Move to 0.

When the result of the determination in step S14 is that the brake switch 14 is in the on state, it is determined that the vehicle is in the braking state, the process proceeds to step S24, and the target cylinder pressure calculation process is performed. It is determined whether or not the target cylinder pressure P ^* _{FL and the} master cylinder pressure P _MC match. If they match, the process proceeds to step S15. If they do not match, the process proceeds to step S25. ..

In this step S25, the variable m is positive.
If m> 0, the above step is directly performed.
Step S17, and if m ≦ 0, step
The process moves to S26. In this step S26, the target series
Pressure P^* _FLAnd current cylinder pressure detection value P _FLError P
_err(= P^* _FL-P_FL) Is calculated and then step S27
Move to.

In this step S27, the variable T _P is calculated according to the following formula 35 in which the value obtained by dividing the error P _err by the reference value P ₀ is rounded off. T _P = INT (P _err / P ₀ ) ... (35) Then, the process proceeds to step S28, the variable m is set to a positive predetermined value m ₀ , and then the process proceeds to step S17.

The processing of FIG. 6 corresponds to the braking force control means. Therefore, when the vehicle is traveling in the non-braking state, the brake switch 14 is in the off state, and therefore the process proceeds from step S14 to steps S15 and S16 to step S17, and T _P > 0. To control signals CS _FL and CS of “0”
_FR is output as a boosting signal to the constant current circuits 20FL and 20FR. Therefore, the constant current circuits 20FL and 20F
No exciting current is output from R, the electromagnetic directional control valves 3FL and 3FR maintain their normal positions, and the front wheel cylinders 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, and each wheel cylinder 1FL
The cylinder pressures of 1 and 1FR are also zero, no braking force is generated, and the non-braking state is continued.

From this state, when the brake pedal 4 is depressed to bring it into a braking state, the process proceeds from step S14 in FIG. 7 to step S24, and the target cylinder pressures P ^* _FL and P calculated in the target cylinder pressure calculation process in FIG. ^* Determine whether or not _FR matches the master cylinder pressure P _MC of the master cylinder 5, respectively. This determination is to determine whether the vehicle is in the straight traveling state or the turning state. In the straight traveling state, as described above, in the processing of FIG. 6, the target cylinder pressures P ^* _FL and P ^* _{FR are set.} Is set equal to the master cylinder pressure P _MC , the steps S24 to S
15, the control signal C as in the non-braking state described above.
By setting both S _FL and CS _FR to zero and setting the electromagnetic directional control valves 3FL and 3FR to the normal position, the master cylinder 5 and the wheel cylinders 1FL and 1FR are brought into communication with each other, and the cylinder pressure P of each wheel cylinder 1FL and 1FR is set. Increase _FL and P _FR to a value equal to the master cylinder pressure P _MC, and set both wheel cylinders 1FL and 1
FR produces equal braking force.

However, when the vehicle is in the braking state in the turning state or when the vehicle shifts to the turning state in the braking state, the target cylinder pressure P ^* _FL (or P ^* _FR ) becomes the master cylinder pressure P in the processing of FIG. 6 described above. Target differential pressure Δ to _MC
Since it is set to a value obtained by subtracting P, in the process for this wheel cylinder 1FL (or 1FR), the process proceeds from step S24 to step S25, and the variable m is set to "0" in the process of the previous step S20. If so, the process proceeds to step S26. Therefore, each target cylinder pressure P ^* _FL (or P ^* _FR ) and the pressure sensor 14FL
An error P _err between the pressure detection value P _FL (or 14 _FR ) and the pressure detection value P _FL (or P _FR ) is calculated (step S26), and this is divided by a set value P ₀ representing an allowable range to calculate a variable T _P (step S26). S27), and then the variable m is set to a predetermined positive value m ₀ , and then the process proceeds to step S17. At this time, when the pressure detection value P _FL (or P _FR ) of each pressure sensor 14FL (or 14FR) does not reach the target cylinder pressure P ^* _FL (or P ^* _FR ), the variable T _P becomes a positive value. Therefore, the process proceeds to step S18, the control signals CS _FL and CS _FR are set to zero, and the pressure increasing mode is continued. When the turning state and the braking state are continued and this flow is repeated, step S1
The variable T _P is decremented by "1" by 9 in step 9 and the variable m is decremented by "1" in step S20. When the variable T _P becomes zero, the process proceeds from step S17 to step S21.
The control signal CS _FL (or CS _FR ) of the predetermined voltage V _S1 is output to the constant current circuit 20FL (or 20FR) as a holding signal. Therefore, the constant current circuit 20FL (or 20FR)
The exciting current corresponding to the predetermined voltage V _S1 from the electromagnetic directional control valve 3
The electromagnetic directional control valve 3FL (or 3FR) is switched to the second switching position by being output to the FL (or 3FR), and the wheel cylinder 1FL (or 1FR) and the master cylinder 5 are disconnected from each other. , A holding mode in which the cylinder pressure P _FL (or P _FR ) of the wheel cylinder 1FL (or 1FR) is maintained at a constant value, and this holding mode is continued until the variable m becomes “0” in step S20.

After that, when the variable m becomes "0", the process proceeds to step S26 again, and when the error pressure P _err becomes less than 1/2 of the set pressure P ₀ at this time, the variable T _P calculated in step S27 is changed. The value becomes "0", the process proceeds from step S17 to step S21 and the above-mentioned holding mode is entered without passing through the pressure increasing mode, and the cylinder pressure P _FL (or P _FR ) of the wheel cylinder 1FL (or 1FR) becomes the target cylinder pressure P ^*. Maintained at _FL (or P ^* _FR ).

Further, each wheel cylinder 1FL (or 1
FR) wheel cylinder pressure P_FL(Or P_FR) Is the goal
Linda pressure P^* _FL(Or P^* _FR) Is higher than
Error P calculated in step S26_errIs a negative value
And the variable T_PAlso becomes a negative value, and from step S18
Up to a predetermined voltage V_S2Control signal CS
_FL(Or CS_FR) Is output as a decompression signal, and
Predetermined voltage V from current circuit 20FL (or 20FR)_S2To
Depending on the exciting current, the electromagnetic directional control valve 3FL (or 3FR)
Is supplied to the third switching position.
It Therefore, the wheel cylinder 1FL (or 1FR)
It is connected to the master cylinder 5 via the hydraulic pump 7F.
That means, wheel cylinder 1FL (or 1FR)
Cylinder pressure P _FL(Or P_FR) Decompression mode
And this is the variable T_PIs maintained until becomes "0"
It

In this way, each wheel cylinder 1F
The cylinder pressures P _FL and P _FR of L and 1 _FR can be made to match the optimum values of the target cylinder pressures P ^* _FL and P ^* _FR by using the cornering power that takes the load movement of the front and rear wheels into consideration, and as a result, the front and rear By using the target yaw rate in consideration of the load movement of the wheels, it is possible to generate a yaw rate that matches the optimum value of the target yaw rate according to the vehicle speed and the steering angle. Therefore, it is possible to prevent the unstable behavior due to the steering in the braking state, improve the steering stability, and improve the transient yaw rate characteristic.

FIG. 8 shows the time-braking characteristics and FIG. 9 shows the time-driving characteristics when the vehicle turns using the braking force control device of this embodiment while turning left while braking. In these figures, the solid line indicates that the braking force is not controlled (control OFF),
The dashed-dotted line shows the braking force controlled without considering the load movement (steady gain fixed), and the broken line shows the braking force control device of the present invention (steady gain variable) with the braking force controlled with consideration of the load movement. 8a shows the braking force difference, FIG. 8b shows the front right wheel braking force, and FIG. 8c shows the front left wheel braking force.
9a shows the generated yaw rate, FIG. 9b shows the acceleration, and FIG.
Shows the speed.

First, as shown in FIG. 8, in the present invention in which the steady gain is variable, the target differential pressure becomes large in the transient state corresponding to the yaw rate rise in FIG. 9a, but the target differential pressure in the steady state is almost "0". It can be seen that the control input, that is, the pressure reduction amount is smaller than that of the fixed gain fixed. As shown in FIG. 9A from the result of FIG. 8, the present invention in which the steady gain is variable has no overshoot of the yaw rate in the initial stage of turning as compared with the case of the control OFF, and FIGS. 9B and 9C.
As shown in (4), it can be seen that a large deceleration occurs and the vehicle speed is sufficiently reduced as compared with the case where the steady gain is fixed.

In the above embodiment, the wheel load sensor 15 which directly detects the wheel load and outputs a detected value corresponding to the wheel load is used as the wheel load detecting means, but the present invention is not limited to this and, for example, a vehicle. The degree of load movement may be estimated from the acceleration / deceleration state of. In the above embodiment, the yaw rate characteristic control is performed by using the method of reducing the master cylinder pressure on one side and reducing the pressure on the other side in order to obtain the target left-right differential pressure, but the present invention is not limited to this. However, the yaw rate characteristic control can be performed by using a traction control actuator, for example, to reduce the pressure on one side and increase the pressure on the other side to generate a differential pressure between the left and right sides.

Further, in the above embodiment, the case where the braking force difference between the left and right wheels on the front wheel side is controlled has been described, but the invention is not limited to this, and the left and right braking force difference between the rear wheels or the front and rear wheels is controlled. You can Further, in the above embodiment, the steering angle sensor 11 is used as the vehicle steering state detecting means.
However, the present invention is not limited to this, and the actual turning angle (actual steering angle) of the wheels may be detected instead of the steering angle sensor. In this case, The steering gear ratio N in Equations 3, 7.6 and 7.7 described above is omitted.

Further, in the above embodiment, the case where the vehicle speed sensor 12 is applied as the speed detecting means has been described, but the present invention is not limited to this, and the vehicle front-rear direction speed is calculated by detecting the wheel speed, the vehicle longitudinal acceleration, and the like. You can also
Still further, in the above embodiment, the braking pressure control device 1
Although the case where the microcomputer is applied as 6 has been described, the present invention is not limited to this, and the electronic circuit such as the comparison circuit, the arithmetic circuit, and the logic circuit may be combined.

In the above embodiment, the feedforward control as shown in FIG. 10 has been described, but the present invention is also effective for the control system using feedback compensation as shown in FIG. 11, and in this case. Is a feedback compensator, for example, detects the generated yaw rate and reduces the error from the target yaw rate by proportional element P, proportional +
It can be configured by an integral element PD, a proportional + integral + differential element PDI, a robust compensator, and the like.

[0065]

As described above, according to the braking force control device of the present invention, the cornering power at the time of braking is calculated from the wheel load detection value of each wheel, and the steady gain of the target yaw rate and the target braking force are thereby calculated. Is calculated, and the braking force control means controls the left and right braking forces on at least one of the front wheels and the rear wheels in accordance with the calculated value. The characteristics can be improved and the steering stability can be improved. Furthermore, since the control input in the steady state, that is, the amount of pressure reduction is reduced, the frequency of operation of the system is reduced, which is advantageous for durability,
The cost can be reduced and the deceleration shortage can be suppressed. In addition, since the steer characteristics in the steady state are close to those in the uncontrolled state, the driver who is accustomed to the uncontrolled vehicle does not feel uncomfortable and wants to positively utilize the characteristics of the uncontrolled vehicle. It is also preferable for the driver.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram showing a basic configuration of the present invention.

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

FIG. 3 is a block diagram showing an example of a braking pressure control device.

FIG. 4 is an explanatory diagram of a vehicle motion model.

FIG. 5 is an explanatory diagram showing the concept of a friction circle.

FIG. 6 is a flowchart showing an example of a processing procedure of a braking pressure control device.

FIG. 7 is a flowchart diagram that follows the flowchart diagram of FIG. 6;

FIG. 8 is a time-braking characteristic diagram of an actual vehicle.

FIG. 9 is a time-traveling characteristic diagram of an actual vehicle.

FIG. 10 is a block diagram of a feedforward control system.

FIG. 11 is a block diagram of a feedback control system.

[Explanation of symbols]

 1FL to 1RR are wheel cylinders (braking means) 2 are actuators 3FL to 3R are electromagnetic directional control valves 4 are brake pedals 5 are master cylinders 11 are steering angle sensors (steering state detecting means) 12 are vehicle speed sensors (speed detecting means) 13 Is a brake switch 14FL to 14MC is a pressure sensor 15F to 15R is a wheel load sensor (wheel load detecting means) 16 is a braking pressure control device

Claims (1)

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, and signals from the steering state detecting means and the speed detecting means. Target yaw rate setting means for inputting and setting a steady gain of the target yaw rate of the vehicle, and at least one of front wheels and rear wheels necessary for realizing the target yaw rate set by the target yaw rate setting means in the vehicle to be controlled. Of the left and right target braking forces of the left and right braking means disposed on at least one of the front wheels and the rear wheels. In a braking force control device, which comprises a braking force control means capable of independently controlling the braking force so as to match the target braking force, a load of each wheel The target yaw rate setting means and the target braking force calculation means are provided with a wheel load detection means for detecting, and at least the steady gain of the target yaw rate is determined according to the cornering power of the wheel estimated from the detection value of the wheel load detection means. A braking force control device which sets and calculates a target braking force.