JPH0550914A - Brake device for vehicle - Google Patents

Abstract

(57) [Abstract] [Purpose] To provide a vehicle brake device in which the rotation of four wheels is suppressed by a braking force of a ratio according to the ground load of each wheel. [Composition] The coefficient of friction μ is such that the actual deceleration G is the target deceleration G _T
If it is larger than 95% and smaller than 105%, it is not changed (S5).
When it is 5% or more, it is increased. Each ground contact load of the four wheels is calculated based on the longitudinal acceleration and the lateral acceleration (S8), the friction coefficient mu, the brake fluid pressure of each wheel on the basis of the target deceleration G _T and the brake factor b is calculated,
The rotation of the wheels is suppressed by being supplied to the wheel cylinder (S9). The braking force required for the entire vehicle is distributed to each wheel at a ratio according to the ground load, the braking ability of each wheel is maximized, and the braking distance is shortened.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle brake device, and more particularly to a control for suppressing wheel rotation based on a ground load of a wheel.

[0002]

2. Description of the Related Art When a vehicle brake is used to suppress the rotation of wheels to decelerate or stop the vehicle, a brake provided on the vehicle is used.
Normally, the ground load of each wheel is different.
It depends on the vehicle structure and load movement based on inertia, etc.When suppressing the rotation of the wheels, the maximum braking force that can be generated differs depending on the magnitude of the ground load, and if the braking force is determined based on the ground load. The braking distance of the vehicle can be shortened. Therefore, in the brake device described in Japanese Patent Laid-Open No. 3-38458, it is necessary for the entire vehicle depending on the ratio of the load applied to the front shaft supporting the left and right front wheels to the load applied to the rear shaft supporting the left and right rear wheels. The braking force is distributed to the front axle and the rear axle, and the rotation of the left and right front wheels and the rotation of the left and right rear wheels are suppressed by the braking forces according to the loads on the front axle and the rear axle, respectively.

[0003]

However, in the distribution of the braking force in the brake device described in the above publication, although the difference between the load on the front wheels and the load on the rear wheels is taken into consideration, the load on the left wheel and the load on the right wheel are taken into consideration. The difference from the wheel load is not considered,
Not enough to stop the vehicle at the shortest distance. An object of the present invention is to provide a vehicle brake device capable of further reducing the braking distance as compared with the conventional one.

[0004]

In order to solve the above-mentioned problems, a vehicle brake device according to the present invention comprises: (a) wheel rotation suppressing means for suppressing each rotation of four wheels; and (b) ) Braking force distribution determining means for distributing the braking force required for the entire vehicle to each of the four wheels at a ratio according to the ground load of each of the wheels, and (c) at least the braking force distribution determining means. And a control unit that controls each of the wheel rotation suppressing units based on the determined braking force ratio.

[0005]

In the brake system according to the present invention, basically, the rotation suppressing means of each of the four wheels is operated by a force having a magnitude corresponding to the ground load of each wheel. As a result, whether the friction coefficient of the road surface is large or small,
All four wheels will be locked almost simultaneously. The braking ability of the four wheels will be extracted most effectively. However, the control means does not necessarily have to control the wheel rotation suppressing means with the braking force ratio determined by the braking force distribution determining means as it is, and the braking force ratio is modified from another viewpoint such as vehicle running stability. In addition to the above, the wheel rotation suppressing means may be controlled.

[0006]

As described above, according to the present invention, the braking ability of the four wheels can be extracted most effectively, and the braking distance can be shortened as compared with the prior art. Further, since the slip ratios of all the four wheels are kept as small as possible, the cornering force is secured and the running stability of the vehicle is increased.

[0007]

Embodiments of the present invention will now be described in detail with reference to the drawings. In FIG. 2, 10 is a brake pedal 10 as a brake operating member. The brake pedal 10 is connected to the master cylinder 12, and hydraulic pressures corresponding to the depression force of the brake pedal 10 are generated in the two pressurizing chambers of the master cylinder 12, respectively. One of the pressurizing chambers of the master cylinder 12 is connected to the front wheel cylinders 26, 2 of the brakes provided on the left and right front wheels 22, 24 by the liquid passages 14, 16 and the branch passages 18, 20, respectively.
8 and the other pressurizing chamber has liquid passages 30, 3
2 and the branch passages 34, 36 allow the left and right rear wheels 38, 4
0 are connected to the rear wheel cylinders 42 and 44 of the brakes respectively installed on the vehicle. 46 is the rear wheel 38, 40
3 is a proportioning valve provided in a liquid passage 32 for use in the.

Electromagnetic directional control valves 50, 52, 54 and 56 are provided in the branch passages 18, 20, 34 and 36, respectively, and hydraulic pressure control valves 58, 60, 62 and 64 are connected thereto. The solenoids of the electromagnetic directional control valves 50-56 are always demagnetized and in the original position shown in the figure, and the wheel cylinders 26, 28, 42, 44 are communicated with the hydraulic control valves 58-64, but the solenoids are energized. If the wheel cylinders 26, 28, 42,
44 is communicated with the master cylinder 12.

The hydraulic pressure control valves 58 to 64 are connected to an accumulator 70 and a reservoir 72 by liquid passages 74 and 76, respectively, and the accumulator 70 has a reservoir 7 therein.
The second liquid is pumped up by the pump 80 and stored in a certain range. The hydraulic pressure control valves 58 to 64 supply the hydraulic pressure of the accumulator 70 to the wheel cylinders 26, 28, 42, 44 by controlling the hydraulic pressure of the solenoid to a height necessary to suppress the rotation of the wheels by controlling the exciting current of the solenoid. Then, the brake operates based on the hydraulic pressure, and the rotation of the wheels is suppressed. In this embodiment, the wheel cylinders 26, 2
A brake (not shown) that operates at 8, 42, 44, the pump 80, the accumulator 70, the hydraulic pressure control valves 58 to 64, and the like constitute wheel rotation suppressing means.

Between the liquid passages 14 and 16 connecting the master cylinder 12 and the front wheel cylinders 26 and 28, and between the master cylinder 12 and the wheel cylinder 4.
Electromagnetic directional control valves 84 and 86 are provided between the liquid passages 30 and 32 connecting the valves 2 and 44, respectively, and the stroke simulators 88 and 90 are connected thereto. The stroke simulators 88, 90 store the brake fluid discharged from the master cylinder 12 to allow the brake pedal 10 to be depressed, and apply a reaction force corresponding to the depression stroke to the brake pedal 10. In the state where the rotation of the wheels is suppressed based on the hydraulic pressure controlled by the hydraulic pressure control valves 58 to 64, the electromagnetic directional control valve 8
The solenoids 4, 86 are demagnetized and the master cylinder 12
Is communicated with the stroke simulators 88 and 90,
For the driver, as if the wheel cylinders 26, 28, 42,
The operation feeling as if it is connected to 44 is provided.

The braking device is controlled by the control device 100. The control device 100 includes a CPU 102 and a ROM 1
04, RAM 106, input unit 108, output unit 110 and a bus. A brake switch 112 for detecting the depression of the brake pedal 10, a pedal force detection device 114 as an operation amount detecting means for detecting the depression force of the brake pedal 10, and a hydraulic pressure of the accumulator 70 are provided in the input unit 108 of the control device 100. The rotational speeds of the hydraulic pressure sensor 116 for detecting, the hydraulic pressure sensors 118, 120, 122, 124 for detecting the hydraulic pressure of the wheel cylinders 26, 28, 42, 44, the left and right front wheels 22, 24 and the rear wheels 38, 40 are detected. Wheel speed sensors 126, 128, 130, 132 for detecting, vehicle height sensors 134, 1 for detecting the height of the vehicle body at each wheel
36, 138, 140 and a front-rear G sensor 144 for detecting the front-rear direction deceleration of the vehicle body are connected.

The front-rear G sensor 144 has a fan-shaped weight supported by a bearing so as to be rotatable around an axis line in the left-right direction of the vehicle. In this weight, a large number of slits are formed at small intervals on an arc centered on the axis of rotation, and two notches are formed at relatively large intervals. By photoelectrically detecting the notch, the origin of the weight, the rotation direction, and the rotation angle are detected, and the longitudinal acceleration is detected.

The output section 110 has hydraulic pressure control valves 58-64.
And electromagnetic directional control valves 50, 52, 54, 56, 84,
86 is connected. In addition, in the ROM 104, FIG.
A map that defines the relationship between the pedaling force of the brake pedal 10 and the target deceleration G _T shown in the graph and the wheel rotation suppression routine shown in the flowchart in FIG. 1 are stored. Hereinafter, suppression of wheel rotation will be described based on this flowchart.

The braking by the hydraulic brake device is usually performed based on the hydraulic pressure controlled by the hydraulic pressure control valves 58-64, and the electromagnetic directional control valves 50-56, 8 are provided.
4,86 are always demagnetized and wheel cylinders 26,2
8, 42 and 44 are communicated with the hydraulic control valves 58 to 64, and the master cylinder 12 has the stroke simulator 8
It is connected to 8, 90. Then, at the same time when the ignition switch is turned on, a main routine (not shown) is executed, the friction coefficient μ of the friction material of the brake is set to the basic value μ _B in the initial setting, and the RAM 10
6 is stored in the friction coefficient storage area. The basic value μ _B is a value determined by design or an actual measurement value of a dry friction material at room temperature.

When the brake pedal 10 is depressed, first, step S1 (hereinafter abbreviated as S1) is executed to read the depression force of the brake pedal 10, the deceleration G of the vehicle body and the vehicle height, and then Target deceleration G at S2
_T is calculated. The target deceleration G _T is calculated from the map that defines the relationship between the stepping force of the brake pedal 10 and the target deceleration G _T.

Next, steps S3 to S7 are executed and the actual deceleration G
The friction coefficient μ of the friction material is determined in accordance with the ratio of to the target deceleration G _T. In S3, it is determined whether the actual deceleration G is 95% or less of the target deceleration G _T , and 9
If it is 5% or less, the determination result of S3 is YES and S6.
Is executed, the friction coefficient μ is determined to be 0.99 μ, and the friction coefficient μ previously stored in the friction coefficient storage area is
Is replaced by Also, the actual deceleration G is the target deceleration G _T
If the actual deceleration G is greater than 95% of the target deceleration G _T , it is determined whether or not the actual deceleration G is 105% or more of the target deceleration G _T. If 105% or more of the target deceleration G _T , then S
The determination of 4 is YES, S7 is executed, and the friction coefficient μ
Is determined to be 1.01 μ. Further, when the actual deceleration G is greater than 95% and less than 105% of the target deceleration G _T , S5 is executed and the friction coefficient μ is determined to the same value as before. As a result, while the actual deceleration G is 95% or less of the target deceleration G _T , the friction coefficient μ is reduced by 1% for each execution cycle of the wheel rotation suppression routine, and 1
While it is more than 05%, it is increased by 1%, and 95% and 10
It will not be changed between 5%.

Next, in S8, the left and right front wheels 22, 2
The wheel load of each of the four wheels and the rear wheels 38, 40 is determined. Four
The magnitude of each load on the wheels depends on the structure of the vehicle and the movement of the load from the rear to the front of the vehicle that occurs during braking, and the same braking force cannot prevent the wheels from rotating so that the four wheels lock simultaneously. The wheel load is determined so that an appropriate braking force can be obtained for each wheel. The load F _FL of the left front wheel 22 is determined according to the following equation. F _FL = W _FL + {(H ・ G _X ) / 2L- (H ・ R _F・
G _Y ) / T} · M However, W _FL : vehicle weight applied to the left front wheel 22 when the vehicle is stopped H: height of the center of gravity of the vehicle G _X : longitudinal acceleration L: wheel base R _F : front wheel roll rigidity distribution G _Y : Lateral acceleration T: Tread M: Vehicle mass

[0018] During braking, the moment of magnitude multiplied by the mass M of the height H and the vehicle center of gravity of the vehicle longitudinal acceleration _{G X (M · H · G} X) occurs, the moment applied from the ground to the front wheels Since the reaction force F is balanced with the moment (F ・ L) obtained by multiplying the wheel base L, F = (M ・ H ・
G _X ) / L is obtained, and this reaction force F is applied to the left and right front wheels 22 and 24. Therefore, the load on the left front wheel 22 is increased by (M · H · G _X ) / 2L. Become.

Further, when the vehicle turns, a load movement occurs in the left-right direction of the vehicle. At the time of turning the vehicle, a moment (M · H · G _Y ) of a magnitude obtained by multiplying the lateral acceleration G _Y by the height H of the center of gravity is generated, and the moment when the reaction force F applied to the left and right front wheels from the ground is applied to the tread T ( Because it is balanced with F ・ T)
= (M · H · G _Y ) / T is obtained. This force F is shared by the front wheels and the rear wheels according to the size of the roll rigidity distribution R _F , R _R. The roll rigidity distribution is a distribution ratio of the front wheel and the rear wheel of the restoring moment transmitted from the suspension device to the sprung weight when the vehicle rotates about the longitudinal axis, (M ・ H ・ G _Y ). A value obtained by multiplying / T by the roll rigidity distribution R _F of the front wheels 22 and 24 is the amount of change in the load on the left front wheel 22 due to turning. If the lateral acceleration G _Y at the time of turning to the left is expressed as a positive value, in the case of the left front wheel 22, the load decreases due to the movement of the load when the vehicle turns to the left.
・ H ・ R _F・ G _Y ) / T is pulled, and when turning right, G _Y becomes a negative value and the load increases.

The load F _{FR on} the right front wheel 24 is calculated by the following equation. F _FR = W _FR + {(H ・ G _X ) / 2L + (H ・ R _F・
G _Y) / T} · M However, W _FR = For vehicle weight right front wheel 24 according to the right front wheel 24 in a stopped state, the load is increased by load transfer laterally during left turn of the vehicle, by adding this The load F _FR is obtained, and the value of G _Y becomes negative when turning right, so the load decreases.

Further, the loads F _RL and F _{RR on} the left rear wheel 38 and the right rear wheel 40 are obtained by the following equations. F _RL = W _RL − {(H · G _X ) / 2L + (H · R _R ·
G _Y ) / T} M F _RR = W _RR − {(H · G _X ) / 2L− (H · R _R ·
G _Y) / T} M However, W _RL: vehicle weight according to the left rear wheel 38 in a stopped state R _R: rear wheel roll stiffness distribution W _RR: front and rear with the vehicle weight brake according to the right rear wheel 40 in a stopped state Since the load on the rear wheel decreases due to the movement of the load in the directional direction, (M · H · G _x ) / 2L is drawn. Further, the moving load in the left-right direction is obtained by multiplying (M ・ H ・ G _Y ) / T by the share ratio R _R of the rear wheel roll rigidity. In the case of the left rear wheel 38, this value is subtracted from the right load. In the case of the rear wheel 40, it will be added.

When the ground loads of the left and right front wheels 22, 24 and the rear wheels 38, 40 are thus obtained, S9 is executed and each wheel is adjusted so that the braking force corresponding to the magnitude of the ground load is obtained. The braking fluid pressures P _FL , P _FR , P _RL , and P _RR supplied to the wheel cylinders 26, 28, 42, and 44 are calculated by the following equations. _{P FL = (F FL · G} T) / (μ · b F) P FR = (F FR · G T) / (μ · b F) P RL = (F RL · G T) / (μ · b R ) P _RR = however _{(F RR · G T) /} (μ · b R), b F front wheel brake factors, b _R is a brake factor of the rear wheels, respectively represented by the following equation. b _F = 2 · A _F · (r / R) b _R = 2 · A _R · (r / R) where A _F : piston cross-sectional area of the brakes of the left and right front wheels 22 and 24 A _R : left and right rear wheels 38, Piston cross-sectional area of 40 brake r: effective radius of disk rotor R: effective radius of tire

Thus, the braking fluid pressures P _FL , P _FR , P _RL , P
_RR is determined based on each ground load of the wheels 22, 24, 38, 40, and by supplying these braking hydraulic pressures P _FL , P _FR , P _RL , P _RR to the wheel cylinders 26, 28, 42, 44. The braking force obtained on each of the wheels 22, 24, 38, 40 is a braking force obtained by distributing the braking force required for the entire vehicle at a ratio according to the ground load of the four wheels 22, 24, 38, 40. It will be.

According to the above equation, the braking fluid pressures P _FL , P _FR ,
By _obtaining P _RL and P _RR , the braking hydraulic pressure P is increased if the actual deceleration G is 95% or less of the target deceleration G _T and the friction coefficient μ is reduced. In this case, since the amount of wheel rotation suppression is insufficient, the braking fluid pressure P is determined to be high, and the amount of wheel rotation suppression is increased. Also, the actual deceleration G is the target deceleration G _T
In the case of 105% or more, since the wheel rotation is excessively suppressed, the friction coefficient is increased, the braking hydraulic pressure P is determined to be low, and the wheel rotation suppression amount is reduced.

Then, in S9, the magnitudes of the exciting currents of the solenoids of the hydraulic pressure control valves 58 to 64 are controlled so that the calculated braking hydraulic pressure is supplied to the wheel cylinders 26, 28, 42 and 44. .. Liquid pressure sensor 118-124
Detected by the wheel cylinders 26, 28, 4
The hydraulic pressure supplied to 2, 44 and the set braking hydraulic pressure P are compared, and the current is feedback-controlled so that the braking hydraulic pressure P is obtained, and the four wheels 22, 24, 3
Braking fluid pressure P set based on each ground load of 8 and 40
Is supplied, the maximum braking force is obtained for each wheel, and the braking distance is shortened.

As described above, in order to obtain the target deceleration G _T corresponding to the stepping force of the brake pedal 10, the height of the braking fluid pressure P is changed depending on the ratio of the actual deceleration G to the target deceleration G _T. While the actual deceleration G is 95% or less of the target deceleration G _T , the friction coefficient μ is decreased by 1% and the braking fluid pressure P is increased each time S6 is executed.
While the actual deceleration G is 105% or more of the target deceleration G _T , the friction coefficient μ is increased by 1% every time S7 is executed, and the braking hydraulic pressure P is decreased. And the actual deceleration G
Is greater than 95% and less than 105% of the target deceleration G _T, the friction coefficient μ is maintained at a constant value, and even if the actual deceleration G does not exactly match the target deceleration G _T , it is within that range. Then, the braking fluid pressure P is kept constant. Therefore, the electromagnetic hydraulic pressure control valves 58 to 64 are not frequently switched between the pressure increasing state and the pressure reducing state, and the rotation of the wheels is suppressed without causing vibration.

As is apparent from the above description, in the present embodiment, the portion for storing the portion for determining the braking fluid pressure P in S9 of the ROM 104 and the portion for executing that portion of the CPU 102 are the braking force distribution determining means. The RO
The control part is constituted by the part for storing the part for performing the feedback control of the braking fluid pressure P in S9 of M104 and the part for executing that part of the CPU 102.

Another embodiment of the present invention is shown in FIGS.
In this embodiment, the actual deceleration G detected by the front-rear G sensor 144 is corrected based on the inclination of the vehicle body with respect to the road surface and the slope of the road surface. Front and rear G sensor 1
Reference numeral 44 denotes a sensor that detects deceleration by detecting a rotation angle and a direction of a weight that rotates around an axis line in the left-right direction of the vehicle. Even when the road surface is inclined and the road surface is inclined, the rotation is performed and an output is obtained. Therefore, front and rear G
The output value of the sensor 144 includes the deceleration, the slope angle of the road surface, and the inclination angle of the vehicle body. When the vehicle body or the road surface has an inclination, the output value is different from the actual deceleration, and this output value is directly used as the target deceleration. If the braking hydraulic pressure is determined in comparison with the speed G _T , the deceleration corresponding to the stepping force of the brake pedal 10 cannot be obtained, so the correction is made.

The road surface slope angle θ is calculated by the road surface slope angle calculation computer 150 shown in FIG. The computer 150 includes wheel speed sensors 126 to 13.
2. The vehicle height sensors 134 to 140, the front-rear G sensor 144, and the pitch rate sensor 152 are connected to each other, and each detection value is input. Pitch rate sensor 15
2 is the same as the yaw rate sensor (described on pages 3-126 to 3-127, May 1991, Toyota Soarer new model car manual) that detects the angular velocity of the vehicle's rotation about the vertical axis using Coriolis force. Is used in a posture in which the sensor axis is parallel to the axis of the vehicle body in the left-right direction, and detects the pitch rate, that is, the rotational angular velocity ω of the vehicle body about its axis in the left-right direction. Also, the computer 15
The ROM of 0 stores the road surface gradient angle calculation routine shown in FIG. 6, and the road surface gradient angle θ is calculated based on this routine and is supplied to the control device 100 for suppressing wheel rotation.

When calculating the road surface slope angle θ, first, S1
21, the wheel speed, front-rear G, vehicle height, and pitch rate sensors 126 to 132, 144, 134 to 140,
The output values of 152 are read respectively. Then S12
2 is executed, and it is determined whether or not the vehicle body speed is constant. The vehicle speed when the determination of S122 is made last time and the vehicle speed when the determination of S122 is made this time are compared, and if the vehicle speed of this time is within the set range with respect to the vehicle speed of the last time, it is constant. It is determined that there is.

If the vehicle body speed is constant, the determination in S122 becomes YES, the count value C of the counter is incremented by 1 in S123, and then, in S124, it is determined whether the count value C is equal to or greater than the set value C _0. , It is determined whether or not the vehicle speed is constant over the set time. The determination in S124 is initially NO, and S129 is executed to calculate the change amount Δθ in the front-rear turning angle of the vehicle body based on the output value of the pitch rate sensor 152. The pitch rate sensor 152 detects the rotational angular velocity ω, and the change amount Δθ is calculated by multiplying one execution cycle time Δt of the road surface gradient angle calculation routine, and Δθ is added to the road surface gradient angle θ in S130. Be done. This θ will be described later.

If the vehicle speed is constant for C ₀ time, S124
Is YES, and the inclination angle θ _A of the vehicle body in the front-rear direction is calculated from the output value of the front-rear G sensor 144 in S125. The deceleration when the vehicle speed is constant is zero, the output value of the longitudinal G sensor 144 for a value corresponding to the inclination angle theta _A of the front and rear direction of the vehicle body, calculating the longitudinal direction inclination angle theta _A The front-rear direction inclination angle θ _A is the sum of the inclination angle of the vehicle body with respect to the road surface and the inclination angle of the road surface. Therefore, next, in S126, after the relative inclination angle θ _B of the vehicle body with respect to the road surface is calculated based on the outputs of the vehicle height sensors 134 to 140, S127
At, the road surface inclination angle θ is obtained by subtracting the relative inclination angle θ _B of the vehicle body from the longitudinal inclination angle θ _A of the vehicle body.

While the vehicle body speed is constant, S122 to S127
Is repeatedly executed, and the road surface gradient angle θ is updated. If the vehicle body speed is not constant, the determination in S122 becomes NO, the count value C is reset in S128, and then the variation amount Δθ in the longitudinal direction of the vehicle body in S129.
Is calculated and added to the road surface gradient angle θ in S130, and this value is set as the road surface gradient angle θ. Deceleration occurs,
If the vehicle body speed is not constant, the output value of the front-rear G sensor 144 will include deceleration, so that S125-
Depending on the execution of S127, the road surface slope angle θ cannot be calculated. Therefore, while the vehicle body speed is not constant, the change amount Δθ in the front-rear direction rotation angle of the vehicle body obtained based on the detection value of the pitch rate sensor 152 is obtained, and is added to the previously obtained road surface inclination angle θ. As a result, the current road surface gradient angle θ is obtained.

Whether the vehicle body speed is constant or not, the change amount Δθ of the vehicle body front-rear rotation angle is obtained from the vehicle body front-rear direction rotation angular velocity ω detected by the pitch rate sensor 152, and the road surface gradient is calculated. Although it is possible to obtain the angle θ, if the road surface gradient angle θ is obtained based only on the output value of the pitch rate sensor 152, the pitch rate sensor 1
The error of the output value of 52 is accumulated and included in the road surface gradient angle θ. On the other hand, as in the present embodiment, while the vehicle body speed is constant, the road surface inclination angle θ is obtained based on the output value of the front-rear G sensor 144 and the vehicle body inclination angle, and when the vehicle body speed is not constant, the pitch rate sensor is obtained. If the road surface gradient angle θ is calculated based on the output value of the 152, the accumulation of the detection error of the pitch rate sensor 152 will be eliminated every time the vehicle speed becomes constant, and the road surface gradient angle can be calculated accurately. You can

In this way, in the road surface gradient angle calculation routine, the road surface gradient angle θ is calculated by executing S125 to S127 while the vehicle body speed is constant, and by executing S129 and S130 when it is not constant. The calculation result is output to the computer of the control device 100.
The wheel rotation suppression routine shown in FIG. 5 is stored in the ROM of the computer of the control device 100, and the stepping force of the brake pedal, the actual deceleration G of the vehicle body, the road gradient angle θ and the vehicle height are read in S101. After the target deceleration G _T is calculated in S102, the actual deceleration G is corrected in S103 based on the road surface inclination angle θ and the vehicle body inclination angle θ _B. The influences of the slope of the road surface and the inclination of the vehicle body with respect to the road surface are eliminated, and S104 to S11
At 0, the braking hydraulic pressure P of the wheel cylinder is determined based on the true actual deceleration G, and the vehicle is accurately braked at the deceleration corresponding to the pedaling force of the brake pedal 10.

Another embodiment of the present invention will be described. In this embodiment, each wheel cylinder 2 of the left and right front wheels 22, 24 is used.
The braking fluid pressures P _FL and P _FR supplied to the valves 6 and 28 are corrected so that the actual yaw rate (hereinafter, referred to as the actual yaw rate) γ follows the target yaw rate γ _M. Also in this embodiment, the braking fluid pressures P _FL , P _FR are equal to the front wheels 22, 24.
The braking fluid pressures P _FL and P _FR thus determined are corrected so as to reduce the difference between the actual yaw rate γ and the target yaw rate γ _M. Although the target yaw rate γ _M is determined by the driver's steering operation, the actual yaw rate γ may deviate from the target yaw rate γ _{M for} some reason such as a difference in the friction coefficient between the left and right road surfaces, resulting in poor maneuverability. Therefore, the actual yaw rate γ
Follows the target yaw rate γ _M so that the vehicle travels as much as the driver intends.

In this embodiment, as shown in FIG.
The control device 100 includes a brake switch 112, a pedaling force detection device 114, a hydraulic pressure sensor 116 for detecting the hydraulic pressure of the accumulator 70, and wheel cylinders 26, 28, 4.
Hydraulic pressure sensors 118 to 124 for detecting the hydraulic pressures of 2, 44,
Wheel speed sensors 126-132, vehicle height sensors 134-14
0 and longitudinal G sensor 144, as well as yaw rate sensor 1
60 is connected. The yaw rate sensor 160 is based on the Toyota Soarer new vehicle manual, May 1991, 3-126.
As described on pages 3-127, the Coriolis force is utilized to detect the angular velocity of rotation of the vehicle about the vertical axis. Further, the wheel rotation is suppressed according to a routine similar to the wheel rotation suppressing routine shown in FIG. 1, but in the present embodiment, the step of determining the braking hydraulic pressure in S9 is the same as the embodiment shown in FIGS. Are different. Less than,
The step of determining the braking hydraulic pressure will be described.

In order to make the actual yaw rate γ follow the target yaw rate γ _M by controlling the braking hydraulic pressures P _FL , P _FR supplied to the wheel cylinders 26, 28 of the left and right front wheels 22, 24, the left front wheel 22 It is sufficient to generate a difference ΔB _Y corresponding to the difference between the target yaw rate γ _M and the actual yaw rate γ between the braking force of B and the braking force of the right front wheel 24. The actual yaw rate γ can be changed by changing the braking force of at least one of the left and right wheels. However, since the amount of change in the yaw rate is almost proportional to the amount of change in the braking force, the braking force difference ΔB _Y is calculated according to the following equation. Be done. ΔB _Y = K _Y (γ _M −γ) where K _Y is a coefficient.

The target yaw rate γ _M is calculated according to the following equation. γ _M = {1 / (1 + T ・ s)} ・ {1 / (1 + K _h・ V)
_{^{2)} · θ / N G}} · V / L However, T: time constant s: Laplace operator K _h: stability factor N _G: steering gear ratio theta: steering angle V: vehicle speed L: wheelbase

When the front wheels 22, 24 rotate by the steering angle α _{0 with} respect to the steering angle θ of the steering wheel, the rotational angular velocity β of the vehicle is expressed by β = (V · α ₀ ) / L. Front wheel 2
When the front wheels 22 and 24 are steered by the steering angle α ₀ when the rear wheels 38 and 40 are non-steered wheels and the rear wheels 38 and 40 are non-steered wheels, the vehicle travels in the traveling direction. The velocity component generated in the direction orthogonal to is approximately V · α ₀ , and the rear wheels 38,
No component in the direction perpendicular to the traveling direction of the vehicle is generated in 40, so L · β = V · α _0, and β = (V · α ₀ ) /
L is obtained. Further, when the steering wheel is steered to turn the vehicle, the steering angle θ is decelerated by the gear train and transmitted to the front wheels 22 and 24. Therefore, if the gear ratio is N _G , α ₀ = θ / N _G.

When the vehicle turns at a very low speed, the turning radius is determined by the steering angle α ₀ , but when turning at a speed V, it is affected by the stability factor K _h . The stability factor K _h is determined by the design specifications of the vehicle, and when the vehicle turns at a speed V, the steering angle required to have the same turning radius as when the vehicle runs at a very low speed is α. For example, the relationship between α and α ₀ is expressed by the following equation. α / α ₀ = 1 + K _h · V ^{2 In} other words, when the actual steering angle is α ₀ , the actual steering angle α is α ₀ / (1 + K _h · V ² ). is there.

Since β = (V · α ₀ ) / L, α
_{Replace 0} with α ₀ / (1 + K _h · V ² ),
Substituting θ / _NG for α ₀ , β = V / L · θ / _NG ·
{1 / (1 + K _h · V ² )} is obtained. A system having the steering angle θ of the steering wheel as an input and the turning angular velocity β of the vehicle as an output can be regarded as a first-order lag system, so β = {1 / (1 + T · s)} · {1 / (1 + K
_h · V ² )} · θ / _NG · V / L is obtained. In the present embodiment, the rotational angular velocity β is set as the target yaw rate γ _M , and the target yaw rate γ _M is obtained by the above equation.

Thus, if the target yaw rate γ _M is obtained and the left and right braking force difference ΔB _Y is obtained from the target yaw rate γ _M and the actual yaw rate γ, the braking force difference ΔB
_The left and right hydraulic pressure difference ΔP _Y required to provide _Y is obtained according to the following equation. ΔP _Y = ΔB _Y / (μ _F · b _F ) Then, the braking fluid pressures P _FL and P _FR are obtained according to the following equations,
P _RL and P _RR are obtained by the same formulas as in the above embodiment. P _FL = F _FL・ G _T / μ ・ b _F + ΔP _Y / 2 P _FR = F _FR・ G _T / μ ・ b _F −ΔP _Y / 2 P _RL = F _RL・ G _T / μ ・ b _R P _RR = F _RR · G _T / μ · b _{R For} the braking fluid pressures P _FL , P _FR of the front wheels 22, 24, see FIG.
As in the embodiment of FIG. 3, the control hydraulic pressure determined according to the ground contact load of the wheel is modified so that the actual yaw rate γ follows the target yaw rate γ _M. The sign of ΔP is + Δ when the counterclockwise yaw rate is positive.
It is determined that P has a positive value.

If the rotation of the wheels is suppressed in accordance with the braking hydraulic pressure thus determined, the actual yaw rate γ is changed to the target yaw rate γ for the braking force of the left front wheel 22 and the braking force of the right front wheel 24.
A difference is caused to follow _M , and the vehicle turns so as to obtain the target yaw rate γ _M.

It should be noted that between the slip ratio of the wheel and the frictional force, the maximum frictional force is obtained when the slip ratio is in a certain range, and until the range is reached, the frictional force is increased as the slip ratio increases. Increases relatively rapidly, and after exceeding that range, the frictional force gradually decreases with increasing slip ratio. Therefore, in the region where the frictional force increases relatively rapidly with the increase of the slip ratio, the hydraulic pressure difference ΔP
It is possible to distribute _Y to the left wheel and the right wheel by ½ to generate the desired yaw rate, but in a region where the frictional force hardly changes as the slip ratio increases,
In the region where the frictional force gradually decreases, even if the braking fluid pressure is increased, the braking force hardly increases, or conversely decreases.

In the region where the frictional force increases relatively rapidly with the increase of the slip ratio, the braking distance can be shortened by increasing the brake operating force. Therefore, the braking fluid can be changed according to the ground load of each wheel. The merit of determining the pressure is that it is possible to secure the cornering force by reducing the slip ratio of each wheel as much as possible rather than shortening the braking distance. In addition, the braking fluid pressure is controlled so that the actual yaw rate follows the target yaw rate. Therefore, the brake device of the present embodiment is used in a region where the friction coefficient of the road surface is higher than the braking fluid pressure. In this case, the braking device can improve running stability.

In the brake system of this embodiment, the effect of shortening the braking distance is truly obtained in the region where the frictional force hardly changes or gradually decreases with the increase of the slip ratio. In, the braking force cannot be increased even if the braking fluid pressure is increased. From this point of view, the distribution of the hydraulic pressure difference ΔP _Y for making the actual yaw rate γ follow the target yaw rate γ _M is not divided into ½ and -3ΔP _Y / 4, respectively. It is recommended that the other allocation ratio be set to + ΔP _Y / 4, or one to −ΔP _Y and the other to 0.

In the above embodiment, the left and right front wheels 2
For 2 and 24, a left and right braking force difference is provided so that the actual yaw rate follows the target yaw rate, but a braking force difference may be provided for the left and right rear wheels 38 and 40, and for all four wheels. A left and right braking force difference may be provided.

Furthermore, in the above embodiment, as shown in FIG.
Similar to the embodiment shown in FIG. 6, the actual deceleration may be corrected based on the road surface inclination angle and the inclination of the vehicle body with respect to the road surface.

In each of the above embodiments,
Although the ground load of the wheel is calculated based on the longitudinal acceleration G _X and the lateral acceleration G _Y , a load detection device may be provided for each wheel to detect the ground load.

Further, in each of the above-mentioned embodiments, the case where the present invention is applied to the electric / manual two-system type brake device has been described as an example, but the present invention is not limited to the brake device having only the electric control system. It can also be applied to other braking devices.

In addition, the present invention can be implemented in various modified and improved modes based on the knowledge of those skilled in the art without departing from the scope of the claims.

[Brief description of drawings]

FIG. 1 is a flowchart showing a wheel rotation suppression routine stored in a ROM of a computer which is a main body of a control device for a vehicle hydraulic brake device according to an embodiment of the present invention.

FIG. 2 is a system diagram of the hydraulic brake device.

FIG. 3 is a graph showing a relationship between a pedaling force of a brake pedal and a target deceleration stored in the ROM.

FIG. 4 is a diagram showing a road surface gradient angle computing computer of a vehicle hydraulic brake device according to another embodiment of the present invention together with a control device for suppressing wheel rotation.

5 is a flowchart showing a wheel rotation suppression routine stored in a ROM of a computer which is a main body of the control device shown in FIG.

FIG. 6 is a ROM of the computer for calculating the road surface slope angle.
5 is a flowchart showing a road surface gradient angle calculation routine stored in FIG.

FIG. 7 is a diagram showing a control device for a vehicle hydraulic brake device that is yet another embodiment of the present invention, together with various sensors that supply detection signals to the control device.

[Explanation of symbols]

 10 brake pedal 22 left front wheel 24 right front wheel 26 front wheel cylinder 28 front wheel cylinder 38 left rear wheel 40 right rear wheel 42 rear wheel cylinder 44 rear wheel cylinder 58 hydraulic control valve 60 hydraulic control valve 62 hydraulic control valve 64 liquid Pressure control valve 70 Accumulator 72 Reservoir 80 Pump 100 Controller 150 Road surface gradient angle computing computer 160 Yaw rate sensor

Claims (1)

[Claims] 1. Wheel rotation suppressing means for suppressing the rotation of each of the four wheels, and a braking force required for the vehicle as a whole depending on the ground load of each of the four wheels. A brake for a vehicle including: a braking force distribution determining unit that distributes in a ratio; and a control unit that controls each of the wheel rotation suppressing units based on at least the braking force ratio determined by the braking force distribution determining unit. apparatus.