JPH03239673A - Car motion control device - Google Patents

Abstract

PURPOSE:To enhance the stability in maneuvering a car by controlling the braking force between the left and right wheels in addition to steering control on the rear wheels, and thereby making the actual yawrate idential to the target yawrate value. CONSTITUTION:The target yawrate is decided by a deciding means 13 on the basis of the car speed and the front wheel steering angle given by sensing means 1, 2, and the actual yawrate currently occurring on the car body is sensed by a sensing means 4, and a rear wheel steering means 5b makes steering of the rear wheels in an amount decided by a steering amount deciding means 5a in the direction that the difference between the two yawrate values is going to nullify. A rotational moment due to difference in the braking force between the left and right wheels is applied in the direction that a braking force giving means 6b eliminates the said difference, and at the same time a braking force is given to at least either of the left and right wheels so that the difference in the braking force between the left and right wheels will become the difference value decoded by a braking force difference deciding means 6a. As a result, the cornering force generated on the rear wheels due to steering on them acts so that the yawrate difference nullifies to gether with the rotational moment generated on the car body, which also acts so that the same difference nullifies.

Description

[Detailed description of the invention] [Industrial application field] The present invention calculates the yaw rate that actually acts on the vehicle body.

The present invention relates to a vehicle motion control device that controls vehicle motion so as to approach a target yaw rate that should be generated in a vehicle body based on the relationship between vehicle motion characteristics.

[Prior Art] Conventionally, this type of device has been disclosed, for example, in
As shown in Publication No. 6, the vehicle body is equipped with a vehicle speed sensor, a front wheel steering angle sensor, and a yaw rate sensor, and a target yaw rate that should be generated in the vehicle body based on the vehicle speed detected by the vehicle speed sensor and the front wheel steering angle detected by the front wheel steering angle sensor. At the same time, a rear wheel steering amount proportional to the difference between the target yaw rate and the actual yaw rate actually acting on the vehicle body detected by the yaw rate sensor is calculated to eliminate the difference between the target yaw rate and the actual yaw rate. By steering the rear wheels by the calculated rear wheel steering amount, cornering force is generated in the rear wheels, and the generated cornering force is used to match the two yaw rates to improve the steering stability of the vehicle. It is said that

[Problem to be solved by the invention '11] However, in general, the cornering force generated on the rear wheels decreases as the rear wheel slip rate increases, the sideslip angle decreases, and the load applied to the rear wheels decreases. There is a tendency to decrease. Therefore, in a vehicle that attempts to match the target yaw rate and the actual yaw rate by generating cornering force on the rear wheels only by steering the rear wheels, as in the conventional device described above, the slip rate of the rear wheels is is high;
When the sideslip angle is small or the load applied to the rear wheels is small, the cornering force cannot be made sufficiently large and it is difficult to match the two yaw rates;
In some cases, it may not be possible to achieve sufficiently good vehicle handling stability.

The present invention has been made to solve the above problem, and its purpose is to control the braking force between the left and right wheels in addition to the steering control of the rear wheels, thereby achieving a match between the target yaw rate and the actual yaw rate. An object of the present invention is to provide a vehicle motion control device that improves the steering stability of the vehicle.

[Means for Solving the Problem] In order to achieve the above object, the structural features of the present invention are as follows:
As shown in FIG. 1, the vehicle motion control device includes a vehicle speed detecting means 1 for detecting the vehicle speed, a front wheel steering angle detecting means 2 for detecting the front wheel steering angle, and a vehicle motion control device based on the detected vehicle speed and front wheel steering angle. A target yaw rate determining means 3 that determines a target yaw rate, a yaw rate detecting means 4 that detects an actual yaw rate occurring in the vehicle body, and a first coefficient is applied to the difference between the determined target yaw rate and the detected actual yaw rate. a steering amount determining means 5a that determines a rear wheel steering amount by multiplying the amount, and a rear wheel steering that steers the rear wheels by the determined rear wheel steering amount in a direction that eliminates the difference between the determined target yaw rate and the detected actual yaw rate. means 5b; braking force difference determining means 6a for determining a braking force difference between the left and right wheels by multiplying the difference between the determined target yaw rate and the detected actual yaw rate by a second coefficient; At least one of the left and right wheels is controlled so that the rotational moment due to the difference in braking force between the left and right wheels acts in a direction that eliminates the difference from the detected actual yaw rate, and the difference in braking force between the left and right wheels becomes the determined braking force difference. a braking force applying means 6b for applying power; a state quantity detecting means 7 for detecting a state quantity of the vehicle that affects the cornering force of the wheels and the barb force; and a steering amount determining means according to the detected state quantity. 5a and a coefficient determining means 8 for determining a second coefficient used by the braking force difference determining means 6a, respectively.

[Operation] In the present invention configured as described above, the target yaw rate determining means 3 determines, based on the vehicle speed detected by the vehicle speed detecting means 1 and the front wheel steering angle detected by the front wheel steering angle detecting means 2, A target yaw rate that is ideal for vehicle running is determined, and the yaw rate detection means 4 detects the actual yaw rate actually acting on the vehicle body, and the rear wheel steering means 5b moves in a direction to eliminate the difference between the two yaw rates. The rear wheels are steered by the rear wheel steering amount determined by the steering amount determining means 5a, and the braking force applying means 6b applies a rotational moment due to the difference in braking force between the left and right wheels in a direction that eliminates the opening of the lid. A braking force is applied to at least one of the left and right wheels so that the braking force difference becomes the braking force difference determined by the braking force difference determining means 6a. As a result, the cornering force generated on the rear wheels by rear wheel steering acts to eliminate the difference between the target yaw rate and the actual yaw rate, and the rotational moment generated on the vehicle body due to the difference in braking force between the left and right wheels is also the same. Therefore, the yaw rate generated in the vehicle body is feedback-controlled by the rear wheel steering and the braking force difference between the left and right wheels so as to reach the target yaw rate.

On the other hand, in determining the rear wheel steering amount and braking force difference, the coefficient determining means 8 determines the state quantities of the vehicle (for example, The first and second coefficients are determined based on the rear wheel slip rate, sideslip angle, load applied to the rear wheels, etc., and the steering amount determining means 5a calculates the first coefficient based on the difference between the target yaw rate and the actual yaw rate. The rear wheel steering amount is determined by multiplying the coefficient, and the braking force difference determining means 6a
The braking force difference between the left and right wheels is determined by multiplying the same difference by a second coefficient. In this case, jll and the second coefficient correspond respectively to the gains of rear wheel steering and feedback control based on the braking force difference between the left and right wheels, so the yaw rate generated in the vehicle body affects the cornering force and split power of the wheels. Feedback control is performed to achieve the target yaw rate by generating cornering force due to rear wheel steering and generating rotational moment due to the difference in braking force between the left and right wheels, with each gain set according to the state quantity of the vehicle that gives the yaw rate.

[Effects of the Invention] As explained above, according to the present invention, the yaw rate generated in the vehicle body is reduced by both the generation of cornering force due to rear wheel steering and the generation of rotational moment due to the difference in braking force between the left and right wheels. Feedback control is performed to match the target yaw rate, so depending on the vehicle condition, such as when the rear wheel slip rate is high, the sideslip angle is small, or the load applied to the rear wheels is small, the rear wheels Even if sufficient cornering force cannot be generated and it is difficult to match the two yaw rates, it is possible to match the two yaw rates by generating a rotational moment due to the difference in braking force between the left and right wheels. Therefore, the handling stability of the vehicle is improved.

In addition, each gain (first and second coefficient) in both feedback controls is expressed as rear wheel slip rate, side slip angle,
Since the determination is made based on the state of the vehicle that affects the cornering force and braking/driving force of the wheels, such as the load applied to the rear wheels, if the condition is such that a large cornering force is generated on the rear wheels. For this purpose, the gain in rear wheel steering control does not have to be too large, and if the rotational moment due to the difference in braking force between the left and right wheels is large, the gain in vJ power difference control cannot be made too large. Even without this, the actual yaw rate can be quickly converged to the target yaw rate. As a result, it is no longer necessary to make each of the feedback gains (first and second coefficients) so large for the purpose of convergence, and both feedback controls can be performed stably.

[Example 1] Hereinafter, an example of the present invention will be described with reference to the drawings.
Figure I2 schematically shows the entire vehicle equipped with a motion control device according to an embodiment of the present invention.

This vehicle includes a front wheel steering device A that steers the left and right front wheels FWI, FW2, a rear wheel steering device B that steers the left and right rear wheels RWI, RW2, and a rear wheel steering device B that steers the left and right rear wheels FWI, FW2, RWI, RW2.
The vehicle includes a braking force applying device C that applies braking force to the vehicle, and an electric control device that electrically controls the rear wheel steering device B and the braking force applying device C.

The front wheel steering device A is displaced in the axial direction to control the left and right front wheels FWl, F.
It has a rack bar 11 for steering W2.

The rack bar 11 is supported within the steering gear box 12 so as to be displaceable in the axial direction.
It is connected to the lower end of the steering shaft 13 inside the steering shaft 13.
It is adapted to be displaced in the axial direction in response to rotation of a steering handle 14 provided at the upper end of the steering wheel. At both ends of the rack bar 11, left and right front wheels FWI are connected via left and right tie rods 15a, left and right tie rods 15b, and left and right knuckle arms 16a and 16b.
FW2 is connected so that it can be steered, and the same front wheels FWI and F
W2 is designed to be steered according to the displacement of the rack bar 11 in the axial direction. steering gear box 1
A control valve (not shown) consisting of a four-way valve is incorporated in the interior of the control valve 2, and the control valve receives an actuation signal supplied from a hydraulic pump 17 via a diversion valve 18 in response to the steering torque acting on the steering shaft 13. Oil is supplied to one oil chamber of the power cylinder 21, and hydraulic oil in the other oil chamber is discharged to the reservoir 22. The power cylinder 21 drives the rack bar 11 in the axial direction in accordance with the supply and discharge of the hydraulic oil to assist in steering the left and right front wheels FWI and FW2.

The rear wheel steering device B is displaced in the axial direction to control the left and right rear wheels RWl, R.
It has a relay rod 31 that steers W2, and the same rod 31
At both ends, as in the case of the front wheel steering device A above,
Left and right rear wheels RWI, RW via left and right tie rods 32a, 32b and left and right knuckle arms 33a, 33b.
2 are connected so that they can be steered. This relay rod 31 is urged to a neutral position by a spring 34 and is driven in the axial direction by a power cylinder 35.

The power cylinder 35 has a spool valve 36 and a lever 3
Together with 7, it constitutes a hydraulic copying mechanism.

The spool valve 36 consists of a valve sleeve 36a fixed to the vehicle body and a valve spool 36b accommodated in the sleeve 36a so as to be slidable in the axial direction. The hydraulic oil supplied from the flow dividing valve 18 is supplied to one oil chamber of the power cylinder 35, and
The hydraulic oil from the other oil chamber of No. 5 is discharged to the reservoir tank 22. The valve spool 36b is connected to one end of a connecting rod 38, and the other end of the rod 38 engages with an intermediate portion of the lever 37 so as to be tiltable and slidable in the axial direction of the lever 37.

The lower end of the lever 37 engages with the lever rod 31 so as to be tiltable and slidable in a direction perpendicular to the rod 31. The upper end of the lever 37 is rotatably connected to the wheel 41 at a position eccentric from the rotation center on the upper surface of the wheel 41. The wheel 41 is engaged with a worm 42 on its outer periphery, and rotates about the rotation center in accordance with the rotation of the worm 42. The worm 42 is connected to the rotating shaft of an electric motor 43, which is a step motor, so as to rotate integrally therewith.

The braking force applying device C includes a braking hydraulic actuator 51. The brake hydraulic actuator 51 consists of four sets of hydraulic actuators for each circumference, and each hydraulic actuator is a three-position switching valve and a pump.

It is a known anti-skid hydraulic actuator or traction hydraulic actuator each configured with a reservoir, an accumulator, etc. (see, for example, Japanese Patent Laid-Open No. 60-248466), and the mask cylinder 5
It receives brake pressure from 2, and maintains and increases pressure from the electric control device.

A control signal indicating pressure reduction is input to set and control the brake pressure in the wheel cylinders 53a to 53d provided for each wheel. The mask cylinder 52 divides the brake pressure that is increased in response to the depression of the brake pedal 54 into front wheel brake pressure and rear wheel brake pressure, and supplies them to the corresponding hydraulic actuators in the brake hydraulic actuator 51, respectively. . The wheel cylinders 53a to 53d are used for left and right front and rear wheels FWI, FW2. RW1. A braking force corresponding to each supplied brake pressure is applied to RW2.

The electric control device includes wheel speed sensors 61a to 61d, load sensors 62a to 62d, oil pressure sensors 63a to 63d,
64 a, 64 b, a front wheel steering angle sensor 65, a yaw rate sensor 66, a lateral acceleration sensor 67, and a rear wheel steering angle sensor 68.

Wheel speed sensors 61a to 61d are for left and right front and rear wheels FW1, FW
2. They are respectively provided in the vicinity of RWI and RW2, and detect the rotational speeds v1 to v4 of each wheel and output detection signals representing the same speeds v1 to V4, respectively.

The load sensors 62a to 62d are front and rear left and right wheels FWI.

FW2. It is installed between RWI, RW2 and the vehicle body, and detects the loads W1 to W4 applied to each wheel, and calculates the same loads W1 to W4.
Detection signals each representing W4 are output.

Oil pressure sensors 63a-63d are wheel cylinders 53a-
53d into each oil passage communicating with each cylinder 53.
Wheel cylinder oil pressure HP in a~53d +~HP
a and outputs detection signals representing the oil pressures HP + to HP a, respectively. The oil pressure sensors 64a and 64b are provided in oil passages communicating with the master cylinder 52, and detect the front and rear wheel mask cylinder oil pressures MP+ and MPr, and detect the same oil pressure M.
Detection signals representing P+ and MP are output. The front wheel steering angle sensor 65 is assembled on the outer periphery of the steering shaft 13, detects the rotation angle of the same shaft 13, and adjusts the left and right front wheels FWI, FW2.
outputs a detection signal representing the steering angle of. The yaw rate sensor 66 is fixed to the vehicle body, and detects the rotational speed of the vehicle body around the vertical axis of the center of gravity to determine the actual yaw rate γ acting on the vehicle body.
Outputs a detection signal representing .

A lateral acceleration sensor 67 is also fixed to the vehicle body, detects the lateral acceleration of the vehicle body, and outputs a detection signal representing the lateral acceleration G. The rear wheel steering angle sensor 68 detects the rotation angle of the electric motor 43 and outputs a detection signal representing the steering angle θr of the left and right rear wheels RWI, RW2 that are steered according to the rotation of the electric motor 43. The front wheel steering angle of and the rear wheel steering angle θr each take a positive value when the left and right front wheels FWI, FW2 and the left and right rear wheels RWI, RW2 are steered to the right, and when each of the wheels is steered to the left, Indicates a negative value. Yaw rate γ also indicates a positive value when it acts on the vehicle body in the clockwise rotation direction, and a negative value when it acts on the vehicle body in the counterclockwise rotation direction. Also, other detected physical quantities v1 to V4? W1~Wa, HP
+~HPa, M Pp, M PR, G are positive values and only their magnitudes are shown.

A microcomputer 71 is connected to each of these sensors. Microcomputer 71 is ROM, RA
It is composed of M, CPU, Ilo (input/output interface), etc., and is pre-stored in ROM. a
rI! By executing the program corresponding to the flowchart of i, the electric motor 43 and brake hydraulic actuator 51 connected to the computer are driven and controlled.

In addition, the ROM of the microcomputer 71 contains a fourth A
Various coefficients of the characteristics shown in FIGS. 7A and 7A and 4B and 7B are stored, and the computer 71
A stability factor setting device 72 is connected to. This stability factor setting device 72 is for selecting and setting the stability factor Kh of the vehicle, and is composed of a selection switch that is selectively operated by the driver.

Next, the operation of the embodiment configured as described above will be explained.

When the ignition switch (not shown) is closed, the microcomputer 71 starts executing the program in step 100, and performs initial setting processing such as clearing various data in the RAM in step 101.
In step 102, a stability factor Kh is set. In setting this stability factor Kh,
The state of the selection switch in the stability factor setting device 72 is read, and the stability factor Kh is set to a value desired by the driver in accordance with the selection state of the selection switch.

Next, in step 103, the left and right front and rear wheel speeds ■1 to V4 indicated by the detection signals from the various sensors 61a to 68
, left and right front and rear wheel loads W, ~W4, left and right front and rear wheel cylinder oil pressure HP +~HP4, mask cylinder oil pressure for front and rear wheels M P r , M P , , front wheel steering angle θf, actual yaw rate γ, lateral acceleration G and rear wheel The steering angle θr is read, and in step 104, the left and right front and rear wheel speeds v1~
The vehicle speed V of the vehicle is calculated using v4. Various methods have been considered for calculating this type of vehicle speed V, but in this example, the peaks (maximum 411) of the wheel speeds ■1 to v4 of each wheel, which change over time, are smoothed. A known method used in anti-skid devices and the like is employed in which the curves connecting the curves are estimated, and the average of each point in time of each of the estimated curves is estimated as the vehicle speed V.

After the process in step 104, in step 105, the stability factor Kh set in the process in step 102, the front wheel steering angle δf input in the process in step 103, and the vehicle speed V calculated in the process in step 104 are used. Then, the target yaw rate γ*, which is the ideal yaw rate that should be generated in the vehicle, is calculated by executing the following calculation formula based on the dynamic characteristics of the vehicle.

7 * = (1/L −, KsL(V/ (1+
K h-V2) LSI In the above calculation formula, Ks are constants representing the wheel base and steering gear ratio specific to the vehicle. After calculating this target yaw rate difference book, in step 106, the difference Δγ = γ tree - γ between the target yaw rate γ* and the actual yaw rate γ inputted by the processing in step 103 is calculated, and the program proceeds to step 1.
The program proceeds to a setting routine for each coefficient consisting of steps 07 to 2.

In step 107, the microcomputer 71
A table provided in the ROM is referred to, and a vehicle speed coefficient aυ(j1
(see figure 4A) and vehicle speed coefficient bu for applying braking force (see figure 4B)
(see figure) are derived respectively.

In step 108, the rear wheel slip rate S is calculated by executing the following calculation formula based on the vehicle speed ■ and the wheel speeds V 3 and Va of the left and right rear wheels RWl and RW2 inputted in the process of step 103. Ru.

S == (2.V -V3-V4)/2 A7 After calculating this rear wheel slip rate Sr, the table provided in the ROM is referred to in step 109, and a table corresponding to the rear wheel slip rate Sr is calculated. Slip coefficient a5 for rear wheel steering
(See Figure 5A) and slip coefficient bs for applying braking force.
(see FIG. 5B) are derived, respectively.

In step 110, the vehicle speed ■ and the left and right front and rear wheel loads W1 to W4 input through the process of step 103 are processed.
The rear wheel load ratio WR is calculated by executing the following arithmetic expression based on .

W R, = (W 3 + W 4 ) / (W +
+W2+W3+Wa) After calculating this rear wheel load ratio WR, the table provided in the ROM is referred to in step 111.

Load coefficient aw for rear wheel steering corresponding to the rear wheel load ratio WRゎ (see Figure 6A), load coefficient b for front wheel braking force application
F1i (see Figure 9J6B) and a load coefficient bRir for applying rear wheel braking force (see Figure 6B) are respectively derived.

In step 2, a table provided in the ROM is referred to, and a lateral acceleration coefficient a for rear wheel steering corresponding to the lateral acceleration G input in the process of step 103 is
G (see Figure 7A) and lateral acceleration coefficient b for applying braking force
G (see 7th BUgi) are respectively derived.

After deriving each of these coefficients, in step 113, the target rear wheel steering angle 01 is calculated.

θr*”-au-as-air”aa”ΔγNext, in step 114, the difference θr*-θr between this target rear wheel steering angle 01 and the detected rear wheel steering angle θr is calculated, and The control signal corresponding to the difference 01 - θr is applied to the electric motor 43 as a control signal representing the amount of rotation of the electric motor 43.
is output to. As a result, the electric motor 43 rotates the left and right rear wheels.

In order to set the steering angle of RW2 at the wheel steering angle θr* with the opening s6, the wheel 41 is rotated via the worm 42 by a rotation amount corresponding to the difference 01-θr.

As the wheel 41 rotates, the upper end of the lever 37, which is provided eccentrically from the center of rotation of the wheel 41, is displaced in the left-right direction in FIG. 2 in accordance with the amount of rotation of the electric motor 43. Due to this displacement, the connecting rod 3 is placed in the middle of the lever 37.
The valve spool 36b assembled via the hydraulic pump 17 is also displaced in the same direction, and a relative displacement occurs between the valve sleeve 36a and the valve spool 36b. The hydraulic oil supplied thereto is supplied to one oil chamber of the power cylinder 35, and the hydraulic oil from the other oil chamber is discharged to the reservoir 22. The power cylinder 35 drives the relay rod 31 by supplying and discharging this hydraulic oil.

In such a case, the spool valve 3 constituting the hydraulic copying mechanism
6 controls supply and discharge of hydraulic oil to and from the power cylinder 35 in cooperation with the relay rod 31 and the lever 37 so as to eliminate relative displacement between the valve sleeve 36a and the valve spool 36b. is displaced in the left-right direction by an amount corresponding to the amount of displacement of the upper end of the lever 41, so the left and right rear wheels RWI, RW2 are steered to the target rear wheel steering angle 01.

In such steering control of the left and right rear wheels RWI and RW2, the direction in which the rear wheels RWI and RW2 are steered and the direction in which the actual yaw rate γ is generated will be described in detail. Now, the left and right front wheels FWI, FW2 are being steered by turning the steering wheel 14, or due to external disturbances such as a crosswind, the target yaw rate γ*
Assuming that a larger actual yaw rate γ occurs in the right rotation direction (or left rotation direction) than the actual yaw rate γ, the yaw rate difference Δγ = γ*−
γ indicates a negative (or positive) value, and the target rear wheel steering angle θr * = -a u'a s-a 1a
a・Δγ indicates a positive (or negative) value, and the left and right rear wheels RWI and RW2 are steered to the right (or left),
Feedback control is performed so that the actual yaw rate γ generated in the vehicle body in the right rotation direction (or left rotation direction) at a rate higher than the target yaw rate γ* matches the target yaw rate γ* due to the cornering force generated at the rear wheels RWI and RW2. be done.

Returning to the explanation of the flowchart in FIG. 3, after the processing in step 114, in step 115, the front wheel braking force difference ΔP+ and the rear wheel braking force difference ΔP are calculated based on the following calculation formula.

ΔPc=-bu-bs-bF1. I-bG・ΔγΔPr
=-bu-bs-bR1bG・Δγ 5.1 In calculating this same braking force difference ΔPr, ΔP, an actual yaw rate γ that is larger in the right rotation direction (or left rotation direction) than the target yaw rate difference is generated. Yaw rate difference Δγ=
If γ*−γ indicates a negative (or positive) value, the braking force differences ΔPr and ΔP each indicate a positive (or negative) value.

After calculating these braking force differences ΔPr and ΔP, step 1
16, the same braking force difference ΔPf, ΔP and the step 1
Based on the mask cylinder oil pressures M P r and M P- for the front and rear wheels inputted in the process of step 03, target oil pressures P1* to P4 for the wheel cylinders 53a to 53d are calculated as follows.

■If the split power differences ΔP+ and ΔP show positive values, the target oil pressures PI* and P3* of the left front and rear wheel cylinders 53a and 53c become M P r+ΔP r and M P−
+ΔP, and the right front and rear wheel cylinders 53b
.

53d target oil pressure P 2 *, P 4 * is MPr,
MP, is maintained.

■If the re-braking force differences ΔPr and ΔP respectively show negative values, the target oil pressures Pl* and P3* of the left front and rear wheel cylinders 53a and 53c are maintained at MPf and MP, and the right front and rear wheel cylinders 53b and 53d target oil pressure P 2 *
, P a books are set to M P t−ΔP+, MP−−ΔP.

In addition, in setting each of the target oil pressures P1* to P4, the target oil pressure of one of the left and right wheel cylinders was always set higher than the mask cylinder oil pressure by the braking force difference ΔPr, ΔPr, but if the brake pedal 54 is depressed and operated. When the oil pressure in each wheel cylinder 53a to 53d is high, the target oil pressure of the left and right wheel cylinders on the opposite side to the above is determined by the braking force difference 1ΔPfl −1 from the mask cylinder oil pressure.
ΔP may be set one minute lower. In addition, the target oil pressure on the increased side is set to 1Δpi/2, 1ΔPi/
The target oil pressure on the side that is 2 minutes richer and lower is 1ΔP "1
/2, ΔP, /2 minutes may be set lower.

After setting ~Pa'I' to each target oil pressure P+', in step 117, the target oil pressure P1*~P4* and each wheel cylinder oil pressure HP+~HPa inputted in the process of step 103 are set. Based on this, the oil pressure in each wheel cylinder 53a to 53d is controlled to become target oil pressure P1* to P4. In this control, each target oil pressure P1~P4* is equal to each wheel cylinder oil pressure HP +
~ HP a, a pressure increase signal is supplied to the hydraulic actuator corresponding to each wheel cylinder 53a to 53d in the braking hydraulic actuator 51, and the hydraulic actuator to which the pressure increase signal is supplied is the wheel cylinder 53a to 53d. Increase brake pressure to. Further, if each target oil pressure P + *~P4 is lower than each wheel cylinder oil pressure HP1~HPa, each wheel cylinder 53a~53 in the braking hydraulic actuator 51
A pressure reduction signal is supplied to the hydraulic actuator corresponding to d, and the hydraulic actuator to which the pressure reduction signal is supplied lowers the brake pressure applied to the wheel cylinders 53a to 53d.

Further, if each target oil pressure P1* to P7 and each wheel cylinder oil pressure HP + to HP a are equal, each wheel cylinder 53 in the brake hydraulic actuator 5
A holding signal is supplied to the hydraulic actuators corresponding to a to 53d, and the hydraulic actuators to which the holding signal is supplied maintain the brake pressure in the wheel cylinders 53a to 53d at the previous state.

As a result, the brake pressures in the wheel cylinders 53a to 53d are set and controlled to the target oil pressures P1 to P4, and the left front and rear wheels FWI, RWI and the right front and rear wheels FW2. Between RW2 and RW2, target oil pressure P1 to PA (braking force difference ΔP+, ΔP
A difference in braking force occurs depending on r). In this case, as mentioned above, the actual yaw rate γ is larger in the clockwise rotation direction (or counterclockwise rotation direction) than the target yaw rate γ*, and the yaw rate difference Δγ=γ*−γ has a negative (or positive) value. If so, the rebraking force differences ΔPf and ΔP each indicate a positive (or negative) value, and the left front and rear wheel cylinders 53a.

Target oil pressure Pl*, P3* of 53c is Δ from target oil pressure P2*, PJ tree of right front and rear wheel cylinders 53b, 53d.
Increased (or decreased) by Pr, ΔP,.

As a result, a rotational moment in the left rotation direction (or right rotation direction) is generated in the vehicle due to the braking force difference, and the actual yaw rate generated in the vehicle body in the right rotation direction (or left rotation direction) is greater than the target yaw rate γ. Feedback control is performed so that γ matches the target yaw rate γ.

After processing step 117, the program executes step 1.
The process returns to step 02, and the W-ring processing consisting of steps 102 to 117 described above continues to be executed, and feedback is performed so that the actual yaw rate γ acting on the vehicle body matches the target yaw rate γ* determined by the dynamic characteristics of the vehicle. remain under control.

As explained above, according to the above embodiment, the left and right rear wheels R
WI, RW2 as target rear wheel steering angle θr*=-au'as'
Steering to au”aa・Δγ, and left and right front wheels FWI, FW
2 and left and right rear wheels RWI, RW2 each braking force difference ΔP "=-
bu-bs-bFlbG・Δγ, ΔP= = −bu
By providing -b s-b R-・ba・Δγ, feedback control is performed so that the actual yaw rate γ matches the target yaw rate γ*, so that the steering stability of the vehicle is improved.

On the other hand, in the feedback control, the coefficient au,
as, av, and aG correspond to gains in rear wheel steering control, and coefficients b u g b s g b F u g b
Since R11g b a corresponds to the gain in the braking force difference imparting control between the left and right wheels, depending on each of these coefficients,
The convergence response of the actual yaw rate γ to the target yaw rate γ is controlled to change. In this case, the coefficients au, bυ
As shown in FIGS. 4A and 4B, when the vehicle speed ■ is large, the value becomes large, so the convergence response becomes faster during high-speed driving when control of the yaw rate is particularly important for the steering stability of the vehicle. This improves the handling stability of the vehicle when driving at high speeds.

In addition, other coefficients as, a11. aa, bs, bpII,
ba, , bG are state quantities of the vehicle that affect the cornering force and the thorn power of the wheels. Coefficient a S, b
As described above, s is determined according to the rear wheel slip rate S (step l○8, 109), and the coefficient as for rear wheel steering control is determined according to the rear wheel slip rate S, as shown in FIG. 5A. The coefficient bs for braking force difference control decreases as the slip ratio Sr increases, and as shown in FIG. 5B, the coefficient bs decreases as the slip ratio Sr increases.
, increases once and then gradually decreases.

Generally, the slip rate of a wheel affects the cornering force and splitting force of the wheel, and as shown in Fig. 8, as the slip rate increases, the cornering force decreases. Moreover, the anti-embryostatic force increases once and then gradually decreases. If we review the rear wheel steering control and braking force difference control of the above embodiment while taking this characteristic into consideration, we will find that when the rear wheel slip rate S is low, the coefficient as (feedback gain) for rear wheel steering control is set to a large value. , and when the same slip rate S is high, the coefficient bs (feedback gain) for braking force difference control is set to a large value. This is done efficiently by complementing each other, and the convergence response side of the yaw rate becomes good.

The coefficients au+bFw and bRw are the rear wheel load ratio W as described above.
It is determined according to R1 (step 110゜11
1) The coefficient a- for controlling the rear wheel steering increases as the rear wheel load ratio WR increases, as shown in FIG. 6A, and the coefficient b Fi for controlling the braking force difference between the front wheels and the rear wheels increases.
r, b R- is the same load ratio W as shown in FIG. 6B.
decrease and increase, respectively, as R increases. In general, the rear wheel load ratio WR affects the cornering force of the rear wheels and the split power of the front and rear wheels, and as the load ratio WR increases, the cornering force of the rear wheels increases. The split power for the front wheels and the rear wheels has a characteristic of decreasing and increasing, respectively. If we review the rear wheel steering control and braking force difference control of the above embodiment while taking this characteristic into account, we can see that when the rear wheel load ratio WR is small, the coefficient bFli (feedback gain) for braking force difference control regarding the front wheels is and when the same load ratio WRゎ is large, both the coefficients a and l (feedback gain) for rear wheel steering si (control) and the coefficient bRu (feedback gain) for braking force difference control regarding the rear wheels are large. Therefore, both of the above-mentioned controls for making the actual yaw rate γ match the target yaw rate γ* are performed efficiently by complementing each other, and the convergence response side of the yaw rate is improved.

The coefficients aG and bG are determined according to the lateral acceleration G as described above (step 2), and the coefficient aG for rear wheel steering control changes as the lateral acceleration G increases, as shown in the jlTA diagram. The coefficient bG for braking force difference control, which gradually increases and then maintains a constant value, is as shown in FIG. 7B.
It decreases as the lateral acceleration G increases. Generally, the lateral acceleration G of a vehicle is approximately proportional to the sideslip angle θ, and the lateral slip angle θ affects the cornering force and braking/driving force of the wheels. As the sideslip angle θ increases, the cornering force has a characteristic of increasing, and the split portion power has a characteristic of decreasing. If we review the rear wheel steering control and braking force difference control of the above embodiment while taking this characteristic into consideration, we will find that when the lateral acceleration G (lateral angle O) is small, the braking force difference control coefficient ba (feedback gain ) is set large and the lateral acceleration G is large, the rear wheel steering control coefficient aG (feedback gain) is set large. The controls complement each other and are performed efficiently, resulting in a good convergence response side of the yaw rate.

In addition, in both feedback control regarding rear wheel steering and braking force application, FIGS. 5A to 5m! As shown in Fig. 7B, the rear wheel steering control is performed according to the efficiency of the vehicle state quantities (rear wheel slip ratio S1, rear wheel load ratio WR, lateral acceleration G) that affect the cornering force and the split power. Each coefficient as, au, aG corresponds to the gain for auxiliary driving force difference control, and each coefficient b corresponds to the gain for sub-driving force difference control.
By setting F 41 g b Rkl p b G, it is not necessary to set each coefficient to a very large value in order to converge the yaw rate. As a result, it becomes possible to suppress each coefficient to a minimum value, and it becomes possible to stably perform both of the above-mentioned feedback controls.

In addition, in the above embodiment, each wheel cylinder 53
Each brake oil pressure within a to 53d is set to target oil pressure P1* to P4.
When setting the actual pressure, the microcomputer 71 compares the target oil pressures P1* to P4* and the detected oil pressures P1 to P4, and controls each hydraulic actuator in the brake hydraulic actuator 51 to perform pressure increase, pressure reduction, and holding. However, each hydraulic actuator is configured as a linear hydraulic actuator, and the microcomputer 71 outputs a control signal representing the target oil pressure P1* to P− to each hydraulic actuator. The linear hydraulic actuator may set each brake oil pressure in each of the wheel cylinders 53a to 53d to target oil pressures P1* to P4 II).

Furthermore, in the above embodiment, the stability factor K
Although h is determined only by operating the selection switch in the stability factor setting device 72, the vehicle speed,
Taking into consideration the driving conditions of the vehicle, such as the coefficient of friction of the road surface,
The stability factor Kh may be determined.

[Brief explanation of drawings]

Fig. 1 is a diagram corresponding to the claims of the present invention described in the above claims, Fig. 2 is an overall schematic diagram of a vehicle equipped with a motion control device showing an embodiment of the present invention, and Fig. 3 is a diagram corresponding to the claims of the present invention. Figures 4A and 4B are characteristic diagrams of changes in vehicle speed coefficients au and bu with respect to vehicle speed V, and Figures 5A and 5B are slip coefficients with respect to rear wheel slip rate S. a s e b
5, Figures 6A and 6B are rear wheel load ratio W.
Figures 7A and IJ7B are characteristic diagrams of changes in load coefficients av, bRii, and bpII with respect to lateral acceleration G, and Figure 8 is a diagram showing changes in lateral acceleration sensors a and b a with respect to lateral acceleration G. Figure 8 is a diagram showing wheel slip ratio S and vehicle It is a change characteristic diagram of the cornering force of the wheel and the movement force of the splinter with respect to the sideslip angle θ. Explanation of symbols A'...Front wheel steering device, B....Rear wheel steering device, C.
・Braking force applying device, D・・Electric control device, FWI,
FW2 ・ ・ ・ Front wheel, RWI, RW
2...Rear wheel, 14...Steering handle, 35...
Power cylinder, 38... Spool valve, 43, Electric motor, 51... Braking hydraulic actuator, 53a
~53d...Wheel cylinder, 61a~61d...
・Wheel speed sensor, 62a to 62d---Load sensor, 8
3a-63d, 64a. 64b... Oil pressure sensor, 65... Front wheel steering angle sensor, 66... Yaw rate sensor, 67... Lateral acceleration sensor, 68... Rear wheel steering angle sensor, 71... Microcomputer.

Claims (1)

[Scope of Claims] Vehicle speed detection means for detecting vehicle speed; front wheel steering angle detection means for detecting a front wheel steering angle; and target yaw rate determining means for determining a target yaw rate based on the detected vehicle speed and front wheel steering angle. , yaw rate detection means for detecting an actual yaw rate occurring in the vehicle body, and steering amount determining means for determining a rear wheel steering amount by multiplying the difference between the determined target yaw rate and the detected actual yaw rate by a first coefficient. and a rear wheel steering unit that steers the rear wheels by the determined rear wheel steering amount in a direction that eliminates the difference between the determined target yaw rate and the detected actual yaw rate, and braking force difference determining means for determining a braking force difference between the left and right wheels by multiplying the difference between the two wheels by a second coefficient; a braking force applying means that applies a braking force to at least one of the left and right wheels so that a rotational moment due to a power difference acts and a braking force difference between the left and right wheels becomes the determined braking force difference; and cornering force and braking of the wheels. a state quantity detection means for detecting a state quantity of the vehicle that affects the driving force; a first coefficient used by the steering amount determination means and a first coefficient used by the braking force difference determination means according to the detected state quantity; 1. A vehicle motion control device comprising: coefficient determining means for determining second coefficients respectively.