JPS6311408A - Active suspension controller - Google Patents

Abstract

PURPOSE:To improve the responsiveness for the steering operation in turning and stability of the car attitude by varying the front/rear wheel distribution ratio of the transfer load between the right and left wheels according to the turning state when the car turns. CONSTITUTION:Each actuator M1 is arranged between a car body B and each wheel W. A turning state detecting means M2 for detecting the turning state including the steering angle of a car is installed. Further, a control means M3 which outputs the signal for controlling the front/rear wheel distribution ratio of the transfer load between the right and left wheels of the car according to the turning state detected by the means M2 is installed. When the car turns, the front/rear wheel distribution ratio of the transfer load between the right and left wheels is varied according to the turning state. Therefore, the turning characteristic of the car is changed according to the actual turning state, and the responsiveness for the steering operation in turning and the stability of the car attitude, etc. are improved.

Description

DETAILED DESCRIPTION OF THE INVENTION Object of the Invention [Field of Industrial Application] The present invention relates to an active suspension control device that is effective when a vehicle turns.

[Prior Art] Conventionally, as a suspension device mounted on, for example, an automobile, an actuator is provided between each wheel of the vehicle and the vehicle body, and a desired displacement amount of each actuator is determined based on the displacement amount and load of each actuator. A so-called active suspension device is known, which independently achieves ride comfort and attitude control by calculating and controlling the actuator.

As an active suspension device as described above, for example, a "vehicle suspension system" (published patent publication No. 60-500662) has been proposed. That is, it includes a wheel suspension device whose displacement can be adjusted, and a means for feedback inputting an electric signal to the device for giving a predetermined displacement according to the load on the suspension device, and the device includes a wheel suspension device whose displacement can be adjusted, and a means for feeding back an electric signal to the device to give a predetermined displacement according to the load on the suspension device. It keeps the vehicle sufficiently stable in all aspects of vehicle operation, regardless of the conditions. Therefore, when a vehicle turns, for example, roll torque is generated by acceleration in the width direction of the vehicle, and hydraulic actuators installed between each wheel and the vehicle body increase or decrease the load acting between each wheel and the vehicle body. By shifting the load between the two, the posture of the vehicle was controlled and ride comfort was improved.

“Problems to be Solved by the Invention” This prior art has the following problems. Namely, because the front and rear wheel distribution ratios of the load transferred between the inner and outer wheels during turning are the same, the vehicle The characteristics of turning performance remained unchanged regardless of the turning state of the vehicle.In general, it is desirable to have oversteer at the start of steering, neutral steer during turning, and understeer when steering is reversed. Therefore, when a vehicle is set to oversteer, its stability decreases due to a sudden change in yaw angular velocity when the steering is reversed.
A vehicle set to have an understeer characteristic has a problem in that there is a delay in the turning state at the start of steering, resulting in a decrease in maneuverability. In general, the characteristics of a vehicle's turning performance are set slightly on the understeer side rather than neutral steer, so
Particularly at the beginning of steering, it is difficult to realize the movement intended by the driver.

SUMMARY OF THE INVENTION An object of the present invention is to provide an active suspension control device that appropriately controls characteristics of turning performance depending on the turning state of a vehicle.

Structure of the Invention [Means for Solving the Problems] The present invention, which has been made to solve the above-mentioned problems, has two wheels each disposed between each wheel of a vehicle and a vehicle body, as illustrated in FIG. an actuator M1, a turning state detecting means M2 for detecting a turning state including a steering angle of the vehicle, and a front and rear wheel distribution of a moving load between left and right wheels of the vehicle according to the turning state detected by the turning state detecting means M2. control means M for outputting a signal for controlling the ratio to the actuator M1;
3, and an active suspension control device characterized by comprising the following.

The actuator M1 changes the load acting between each wheel and the vehicle body, for example. For example, the hydraulic actuator may include a hydraulic actuator including a piston and a cylinder, a hydraulic power source, and a servo valve that communicates or shuts off the hydraulic power source and the hydraulic actuator. In this case, the load can be changed in accordance with a change in the displacement amount of the piston of the hydraulic actuator.

The turning state detection means M2 detects the turning state of the vehicle, including at least the steering angle. For example, steering angle,
It can be composed of sensors that detect vehicle speed, yaw angular velocity, etc. For example, a sensor that detects acceleration in the vehicle width direction, a sensor that detects longitudinal acceleration, a sensor that detects the load acting between each wheel and the vehicle body, or a combination of each of the above sensors and an acceleration sensor may be used. Good too. As the acceleration sensor, for example, a strain gauge may be used, or a servo accelerometer may be used. In addition, as a yaw angular velocity sensor, for example, a rate gyro,
A photographing gyro, a fiber optic gyro, etc. can be used. Further, as the steering angle sensor, for example, a well-known rotary encoder capable of detecting the origin position may be used.

The control means M3 outputs a signal for controlling the front and rear wheel distribution ratio of the moving load between the left and right wheels of the vehicle, depending on the turning state of the vehicle. For example, it is possible to control the front wheel distribution ratio of the moving load to be small when steering is started, and to control the rear wheel distribution ratio of the moving load to be small when the steering is reversed. Further, for example, the front/rear wheel distribution ratio of the moving load may be controlled so that the turning state becomes the target turning state.

Here, the turning state becoming the target turning state may mean, for example, that the yaw angular velocity of the vehicle approaches the target yaw angular velocity calculated from the steering angle and the vehicle speed. Such control is, for example, based on a value obtained by multiplying the deviation between the detected yaw angular velocity and the target yaw angular velocity calculated from the detected steering angle and vehicle speed by the detected yaw angular velocity. This can be achieved by controlling the load distribution ratio between the front and rear wheels. For example, based on the value obtained by multiplying the deviation between the detected yaw angular velocity and the target yaw angular velocity calculated from the detected steering angle and vehicle speed by the detected vehicle width direction acceleration,
The front and rear wheel distribution ratio may be controlled. Furthermore, for example, the detected steering angle, the vehicle speed, the square value of the vehicle speed, and the deviation between the detected yaw angular velocity and the target yaw angular velocity calculated from the understeer setting coefficient predetermined based on the vehicle characteristics. Based on the value multiplied by the yaw angular velocity,
The front and rear wheel distribution ratio may be controlled. Note that, for example, if the target yaw angular velocity calculated above exceeds the allowable yaw angular velocity determined from the vehicle width direction acceleration and the vehicle speed at the time of turning, which are predetermined based on the vehicle characteristics, the allowable yaw angular velocity is set as the target yaw angular velocity. It may be configured to have an angular velocity. Here, the allowable yaw angular velocity can be determined, for example, by dividing a predetermined acceleration in the vehicle width direction determined according to the coefficient of friction between the wheels and the road surface by the vehicle speed at the time of turning. The control means M3 can be realized, for example, by an independent discrete logic circuit. Alternatively, for example, the circuit may be configured as a logic operation circuit together with a well-known CPU, ROM, RAM, and other peripheral circuit elements, and may perform operations and output signals according to a predetermined processing procedure.

"Operation" As illustrated in FIG. 1, in the active suspension control device of the present invention, the control means M3 controls the amount of load transferred between the left and right wheels of the vehicle depending on the turning state detected by the turning state detecting means M2. The signal that controls the distribution ratio is set to 7 units.
It works to output to 1.

That is, when the vehicle turns, the front and rear wheel distribution ratio of the load transferred between the left and right wheels is changed depending on the turning state. Here, the load on each wheel during turning and cornering power have a nonlinear relationship as shown in FIG. 2. As shown by arrow a in the figure, the cornering power when the moving load between the left and right (inner and outer) wheels is small is the sum of the value CP2I on the inner wheel side and the value CP2O on the outer wheel side. On the other hand, as shown by the arrows in the figure, the cornering power when the moving load between the left and right (inner and outer) wheels is large is the sum of the value CP11 on the inner wheel side and the value CP10 on the outer wheel side. The magnitude relationship between the above two phases is determined as shown in the following equation (1).

CPI I+CP10<CR2I+CP2O...(1
) In this way, the smaller the moving load between the left and right (inner and outer) wheels during turning, the greater the cornering power. Further, the characteristics of the turning performance of the vehicle are determined based on the value of the following equation (2).

CRxLR-CFxLF - (2> However, CR......Rear wheel side cornering power LR.
...Distance CF between rear wheel axle and vehicle center of gravity...
Front wheel side cornering power LF... Distance between the front wheel axle and the vehicle center of gravity Here, if the value of formula (2) is negative, it is oversteer, if it is O, it is neutral steer, and if it is positive, it is understeer. For this reason, when turning left and right (inside and out)
When controlling the front and rear wheel distribution ratio of the inter-wheel transfer load, reducing the front wheel distribution ratio increases the cornering power on the front wheels, resulting in oversteer characteristics, while decreasing the rear wheel distribution ratio increases the cornering power on the rear wheels. This results in understeer characteristics.

Therefore, the active suspension control device of the present invention has the following features:
When the vehicle turns, it works to change the characteristics of turning performance according to the turning state. The technical problems of the present invention are solved by the functions of the main components of the present invention as described above.

[Example] Next, a preferred example of the present invention will be described in detail based on the drawings. FIG. 3 shows the system configuration of the first embodiment of the present invention.

In the figure, a vehicle 1 has a suspension 5.6 between the vehicle body 2 and left and right front wheels 3.4, and a suspension 9.10 between the vehicle body 2 and left and right rear wheels 7 and 8. . Each suspension 5, 6, 9.1O has a displacement converter 11, 12, which outputs an analog signal proportional to its displacement.
13, 14, a load sensor 1 consisting of a load cell that measures the load acting between each wheel 3, 4, 7, 8 and the vehicle body 2;
5゜16.17, 18, unsprung acceleration sensor 19 arranged on each suspension arm to detect unsprung acceleration
, 20, 21, 22. and each suspension 5, 6,
9. Servo valves 23, 24, which adjust the displacement amount of 10.
25 and 26 are provided, respectively.

Additionally, a vehicle speed sensor 27 that detects the vehicle speed of the vehicle 1, a steering angle sensor 28 that detects the steering angle, a longitudinal acceleration sensor 29 that is disposed near the center of gravity of the vehicle 1 and detects acceleration in the longitudinal direction, and a longitudinal acceleration sensor 29 that detects acceleration in the longitudinal direction. A vehicle width direction acceleration sensor 30 that detects acceleration and a yaw rate sensor 31 that detects yaw rate are also provided.

The detection signals of each of the above sensors are transmitted by the electronic control unit (hereinafter simply EC).
The ECU 40 drives each servo valve 23, 24, 25.26 to control each suspension 5, 6, 9.10.

Since the configurations of the suspensions 5, 6, 9, and 10 are all the same, the left front wheel suspension 5 will be explained based on FIG. 4 as an example. The left front wheel suspension 5 supports the left front wheel 3 at the other end of a suspension arm 51, one end of which is rotatably attached to the vehicle body 2. A coil spring 53 and a hydraulic actuator 54 housed inside the coil spring 53 are arranged in parallel between the lower end of the support member 52 whose upper end is rotatably attached to the vehicle body 2 and the suspension arm 51. There is. The hydraulic actuator 54 includes a cylinder 55 and a cylinder 5.
5 Piston 5 that separates the interior into an upper chamber 56 and a lower chamber 57
The lower end of a rod 59 extending downward from the piston 58 is rotatably attached to the suspension arm 51.

The load applied to the hydraulic actuator 54, that is, the load acting between the vehicle body 2 and the left front wheel 3, is measured by the left front wheel load sensor 15, which is a load cell disposed inside the support member 52.

Further, the displacement amount of the piston 58 is measured by a left front wheel displacement amount converter 11 whose one end is attached to the suspension arm 51 and the narrow end is attached to the support member 52. Further, the unsprung acceleration is detected by the left front wheel unsprung acceleration sensor 19 disposed near the end of the suspension arm 51 that supports the left front wheel 3 .

The upper chamber 56 and lower chamber 57 of the hydraulic actuator 54 are connected to the electromagnetic front left wheel servo valve 23 through conduits 60, 61, respectively. The left front wheel servo valve 23 is
A hydraulic circuit includes a reservoir 62 and a pump 63. High-pressure hydraulic oil boosted by the pump 63 is constantly supplied to the left front wheel servo valve 23 , which passes the hydraulic oil through a variable orifice inside thereof, and then transfers the hydraulic oil to the reservoir 62 . return. The left front wheel servo valve 23 can control the pressure difference in the internal pressure between the upper chamber 56 and the lower chamber 57 of the hydraulic actuator 54 to an arbitrary value by adjusting the flow rate of the hydraulic oil using the variable orifice. . Therefore, ECU40
When the left front wheel servo valve 23 is driven and controlled, the piston 58 of the hydraulic actuator 54 is displaced due to the pressure difference, and the load acting between the vehicle body 2 and the left front wheel 3 is adjusted.

Next, the configuration of the ECU 40 mentioned above will be explained based on FIG. 5. The ECU 40 includes a CPU 40a, a ROM 40b,
It is configured as a logic operation circuit centered around RAM40c, etc.
Input/output ports 4Qe, 40f via common bus 40d
is connected to perform input/output with the outside.

The ECU 40 includes a signal adjustment circuit 40q equipped with a buffer or filter for the detection signals of each sensor described above, a multiplexer 40h that selectively inputs each detection signal, and an A/D converter 40i that converts analog signals into digital signals.
These detection signals are input to the CPU 40a via the input/output port 40e.

Further, the ECU 40 controls each servo valve 23, 24°, 25.
26 drive circuits 40j, 40, 40m.

40n and a D/A converter 40p that converts digital signals into analog signals, the CPU 40a connects the drive circuits 40j and 40 to each of the above drive circuits 40j and 40 via an input/output port 40f.
A control signal is output to m and 40n.

Next, the relationship between various quantities used for control in the first embodiment will be explained based on FIG. 6.

The various quantities detected by each of the sensors described above are listed below. That is, the displacement amount of the suspension arranged for each wheel 3, 4, 7.8 X1°X2. X3. X4,
Load f1. f2. f3. f4 and unsprung acceleration XLJ
1. Xu2. Xu3. Xu4 is the displacement converter 11,
12, 13, 14, load sensor 15, 16.17.18
and detected by unsprung acceleration sensors 19, 20, 21, and 22. Further, longitudinal acceleration XCC7, vehicle width direction acceleration YCQ, and yaw rate (yaw angular velocity) γ acting on the center of gravity G of the vehicle are detected by longitudinal acceleration sensor 29, vehicle width direction acceleration sensor 30, and yaw rate sensor 31. Furthermore, the vehicle speed V and steering angle θ are detected by a vehicle speed sensor 27 and a steering angle sensor 28.

Based on these quantities, the motion state of the suspension arranged corresponding to each wheel 3, 4, 7.8 is calculated based on the center of gravity G of the vehicle.
into four types of motion states. In other words, the heap (Heave) is the vertical vibration shown by the arrow H of the center of gravity G.
, pitch (pitch), which is a longitudinal vibration indicated by an arrow P around an axis in the vehicle width direction passing through the center of gravity G; roll (roll), which is rotation around an axis in the longitudinal direction indicated by an arrow R around an axis in the longitudinal direction passing through the center of gravity G; Warp (Warp > 4
It is a kind of movement state.

Next, a target displacement amount of the center of gravity G is calculated from the four types of movement states described above, corresponding to each movement state. That is, there are four types: heap target displacement amount 1-1d, pitch target displacement amount pd, roll target displacement amount Rd, and warp target displacement amount Wd. Furthermore, the four types of target displacement amounts of the center of gravity G are set for each wheel 3, 4.
, 7.8, the target displacement amount Xd1. of each suspension provided corresponding to 7.8. Xd2. Xd3. Convert to Xd4. E
The CU 40 controls each servo valve so that the displacement amount of each suspension becomes the target displacement amount. The wheelbase of the vehicle is such that the distance between the center of gravity G of the L1 vehicle and the front wheel axle is Xf, the front wheel tread is lf, and the rear wheel tread is Tr.

Next, suspension control processing executed by the ECU 40 will be explained based on the flowcharts shown in FIGS. 7(A) and 7(B). This suspension control process is performed by EC
After U40 is activated, it is repeatedly executed during each activation waiting period.

First, in step 100, initialization processing such as clearing the RAM 40c and setting initial values is performed.

In the subsequent step 110, a process is performed in which the values obtained by A/D converting the detection signals of each of the sensors described above are read. That is, the displacement iX1. X2. X3. X4, load f1. f2
．． f3. f4 and unsprung acceleration Xu1. Xu2. Xu
3. Xu4, longitudinal acceleration XcΩ, vehicle width acceleration Y
The values of Cg, vehicle speed V, steering angle θ, and yaw rate γ are read.

Next, the process proceeds to step 120, and as described above, the displacement of each suspension detected this time! X1°X2. X3. X
Based on 4, the current heap displacement amount XHn at the center of gravity
, pitch displacement amount XPn, roll displacement amount XRn, and warp displacement amount XWn are calculated as shown in the following equations (3) to (6).

XHn=X1+X2+X3+X4 −(3)XP
n- ((X1+X2) -(X3+X4))xAPl-(4
)XRn= (Xi-X2)XAR1 + (X3-X4)XAR2...(5)XWn= (
Xi-X2)xARl-(X3-X4)XAR2...(6) However, API=
1/L AR1= (Xf/L)x (1/Tf)AR2= (
(L-Xf)/L) ) from heap speed DXH, pitch speed D
XP, roll speed DXR, and warp speed DXW are calculated using the following formula (7
) to (10) are performed. Here, D represents a time differential operator d/dt.

DXH=XHn-XHn-1-(7) DXP=XPn-XPn-1,-(8)DXR=XRn
−XRn−1・ (9)DXW=XWn−XWn−1・
(10) Next, proceed to step 140, load F1 acting between the left front wheel 3 and the vehicle body 2, load F2 acting between the right front wheel 4 and the vehicle body 2, and load F2 acting between the left rear wheel 7 and the vehicle body 2. load F3 acting on the right rear wheel 8 and load F3 acting between the right rear wheel 8 and the vehicle body 2
4 is calculated as shown in the following equations (11) to (14). Here, the mass Mf is obtained by subtracting the front wheel pseudo mass mf, which is an arbitrary value set in advance, from the front wheel unsprung mass Mfd, and the mass Mr is an arbitrary value, which is set in advance, from the rear wheel unsprung mass Mrd. It is obtained by subtracting a certain rear wheel pseudo mass mr.

F1=f1−MfxXu 1=(
11) F2=f2-MfxXu2-(
12) F3=f3-MrxXu3-(
13) F4=f4-MrxXu4・
(14) In the following step 150, the above step 140
Based on each load F1~'4 calculated in , the heap load FH1 pitch torque FP10-le torque F
A process is performed to calculate R and warp torque FW as shown in the following equations (15) to (18). FH=F1+F2+F3
+F4...(15)FP= (F1+F2
)xAP3 - (F3+F4)xAP4...(16)FR
= (Fl-F2)XAR3 + (F3-F4)xAR4...(17)FW= (
Fl-F2)XAR3-(F3-F4)XAR4...(18) However, AP3
=Xf AP4=L-Xf AR3=Tf/2 AR4=Tr/2 Next, the process proceeds to step 160, where each speed DXH, DXP, DXR calculated in step 130 above and step 15 above are calculated.
Based on the load and torques FH, FP, and FR calculated at 0, the heap target displacement amount 1-1d at the center of gravity position,
The pitch target displacement amount pd and the roll target displacement amount Rd are calculated using the following formula (
19) to (21) are performed.

Hd=AH5XFH-AH6XDXH...(19)P
d=AP5xFP-AP6xDXP +AP7xXcg...(20)Rd=
AR5xFR-AR6xDXR +AR7XYCC1...(21) However, the stiffness in the heap direction is KH, and the damping coefficient in the heap direction is C.
When H, AH5=1/KH AH6=CH/KH.

In addition, the pitch direction stiffness is KP, and the pitch direction damping coefficient is C.
When P, AP5=1/KP and AP6=CP/KP.

AP7 is a longitudinal acceleration correction coefficient.

Furthermore, if the roll direction rigidity is KR and the roll direction damping coefficient is CR, then AR5=1/KR AR6=CR/KR AR7 is the vehicle width direction acceleration correction coefficient.

In the subsequent step 170, a target yaw rate is determined from the vehicle speed V and the steering angle θ, and the deviation between the target yaw rate and the actually detected yaw rate γ is further multiplied by the yaw rate γ to calculate a yaw rate correction coefficient Y. Warp speed DXW calculated in step 130 and step 15 above
From the warp torque FW calculated at 0 and the yaw rate correction coefficient Y, a process is performed to calculate the warp target displacement IWd at the center of gravity position as shown in the following equations (22) and (23).

Y= (AWIXVXθ-γ)×γ0AW2...(2
2) Wd = AW5xFW - AW6xDXW + AW7xYcgxY ... (23) However, AWl is the yaw rate gain, AW2 is the front-rear load distribution correction coefficient when traveling straight, and the warp direction stiffness is KW,
If the warp direction damping coefficient is CW, then AW5=1/KW AW6=CW/KW.

Note that AW7 is a warp correction coefficient.

By the way, the stiffness KH, KP, KR in each of the above directions.

KW and the directional damping coefficients CH, CP, OR, and CW are each preset arbitrary values.

Next, proceed to step 180 and proceed to step 160.17 above.
Each target displacement amount Hd, Pd at the center of gravity position calculated at 0
, Rd, Wd, each target displacement amount Xd1. Xd2. Xd3. Xd4 is expressed by the following equations (24) to (2
The calculation process shown in 7) is performed.

Xd1= (1/4)X((Hd+AP8xPd)+(
AR8xRd+AR9xWd))...(24)Xd2
= (1/4)X((Hd+AP8XPd)-(AR8
XRd+AR9xWd))...(25)Xd3-(1
/4)X((Hd-AP8xPd)+(AR8xRd-
AR9xWd))...(26)Xd4= (1/4)
X((Hd-AP8xPd)-(AR8XRd-AR9
XWd))...(27) However, AP8=1= (1/
AP1 > AR8-(LxTf)/Xf= (1/AR
1>AR9= (LxTr)/ (L-Xf>-(1/
AR2> In the following step 190, each target displacement amount Xd1. Xd2. Xd3°Xd4 and the currently detected saleta displacement amount Xi, X2. X3゜Deviation from X4 XO
1, XO2, XO3, and XO4 using the following formulas (28) to (31)
The calculation process is performed as follows.

XOl = A1 X (Xd 1-Xl ) +A10
...(28)XO2=A2X (Xd2-X2)+A
20. (29)XO3=A3x (Xd3-X3)+
A30・ (30)XO4=A4x (Xd4-X4)
+A40...(31> However, A1.A2.A3.A4
is a gain constant determined according to the responsiveness of each hydraulic actuator as described above -26 = number, AIO, A20. A30. A40 is an offset constant determined according to the target standard vehicle height.

Next, the process proceeded to step 195, and voltages corresponding to the deviations XO1, XO2, XO3, and XO4 calculated in step 190 were output to each servo valve 23, 24, 25, and 26 of each suspension 5, 6, and 9° 10. After that, the process returns to step 110 described above. Thereafter, this suspension control process is executed by repeating steps 110 to 195.

Next, an example of the above control will be explained with reference to the timing chart of FIG. 8. This figure shows changes in various quantities over time when the vehicle is running in a slalom with the left and right steering angles being the same, with the vehicle speed V being constant.

During time To-T1, since the steering angle is small and the yaw rate deviation is small, it is considered that the vehicle is not yet in a turning state, and the front and rear wheel distribution ratios of the moving load between the inner and outer wheels are set to be the same. Therefore, the turning performance of the vehicle remains the same as the preset basic turning characteristics.

At time T1, the yaw rate of the vehicle lags behind the target yaw rate. Therefore, the front wheel side of the front wheel distribution ratio of the transfer load between the inner and outer wheels is made smaller to increase the cornering power on the front wheel side.

As a result, the turning performance of the vehicle is changed from the basic turning characteristics to oversteer. Therefore, the responsiveness of the vehicle to the steering angle θ is improved, and the yaw rate approaches the target yaw rate, as shown by the broken line in the figure. At time T2, the yaw rate delay is corrected and the yaw rate deviation disappears. Therefore, at the same time T2, the front and rear wheel distribution ratios are made the same, and the turning performance of the vehicle is returned to the basic turning characteristic. In addition,
If the turning performance of the vehicle was not changed as described above, the yaw rate would be delayed and the maneuverability of the vehicle would be reduced, as shown by the dashed line in the figure.

At time T3, a steering reversal is performed. Therefore, the front and rear wheel distribution ratio of the load transferred between the inner and outer wheels is reduced on the rear wheel side to increase the cornering power on the rear wheel side. This results in
The turning performance of the vehicle is changed from the basic turning characteristics to the understeer side. Therefore, the vehicle does not excessively follow the start of a turn due to steering reversal, and the vehicle transitions to stable turning without causing any entanglement or spin. Note that, if the turning performance of the vehicle was not changed as described above, the yaw rate would advance as shown by the dashed line in the figure, and the stability of the vehicle would decrease.

At time T4 after the vehicle enters a stable turning state,
By making the front and rear wheel distribution ratios the same again, the turning performance of the vehicle is returned to the basic turning characteristics. Thereafter, the front wheel side of the front and rear wheel distribution ratio is reduced again, and the turning performance of the vehicle is changed from the basic turning characteristic to the side of oversteer. As a result, at time T5, the yaw rate matches the target yaw rate, and the yaw rate deviation disappears. Therefore, at the same time T5, the front and rear wheel distribution ratios are made the same, and the turning performance of the vehicle is returned to the basic turning characteristic. Thereafter, control is continued to change the turning performance of the vehicle by adjusting the front and rear wheel distribution ratio in accordance with changes in the steering angle of the vehicle.

In this first embodiment, the suspension is 5°6.9
．． 10 corresponds to the actuator M1, and the vehicle speed sensor 27, the steering angle sensor 28, and the yaw rate sensor 31 correspond to the turning state detection means M2. In addition, ECU40 and the E
Processing executed by CU 40 (step 160.17
0, 180, 190, '195) functions as the control means M3.

As explained above, in the first embodiment, the rigidity and damping coefficient in four directions are set in advance, and the
Target displacements IHd and Pd for various motion states: heap, pitch, roll, and warp.

Calculate Rd and Wd, and apply the calculated values to each suspension 5,
6, 9.10 target displacement amount Xdl, Xd2°Xd3. X
d4, and when controlling each suspension 5, 6, 9.10 according to the target displacement amount, the target yaw rate and the yaw rate are set to the target yaw rate calculated based on the vehicle speed V and the steering angle θ. The deviation from the yaw rate is further multiplied by the yaw rate to calculate a yaw rate correction coefficient Y according to the turning state, and the warp target displacement &iWd is determined based on the yaw rate correction coefficient Y. Therefore, at the start of steering, the front and rear wheel distribution ratio of the load transferred between the left and right (inner and outer) wheels becomes smaller on the front wheels, and on the other hand, when the steering is reversed, the front and rear wheel distribution ratio becomes smaller on the rear wheels. Therefore, the turning performance of the vehicle changes from the basic turning characteristics (for example, near neutral steering) to the oversteer side at the start of steering, once changes to the understeer side when the steering is reversed, and then returns to the original basic turning characteristics again. This makes it possible to start turning quickly in accordance with the driver's intention at the start of steering, and to maintain a stable turning state when the steering is reversed. In this way, the turning performance changes suitably depending on the steering condition.

In addition, as the turning performance of the vehicle changes depending on the steering angle as described above, there is no need for a complicated configuration such as the so-called four-wheel steering that steers all the way to the rear wheels. By simply controlling the suspension by changing the front and rear wheel distribution ratio, turning performance comparable to four-wheel steering can be achieved.

Furthermore, when calculating the yaw rate correction coefficient Y, a yaw rate correction value is used, which is the deviation between the target yaw rate and the yaw rate multiplied by the yaw rate. Characteristics, understeer side, and basic turning characteristics change smoothly in that order. This improves the control accuracy of the vehicle during steering, and increases the responsiveness and stability of the vehicle.

There are also four types of motion states. The target displacement of the suspension is calculated based on heap, pitch, roll, and warp, and the warp target displacement takes into account the torsion of the vehicle body, increasing the degree of freedom in control and stabilizing the vehicle posture. can be done.

Furthermore, so-called active control of the suspension becomes possible, suppressing changes in vehicle posture, and improving maneuverability during turns, making it possible to achieve both maneuverability, stability, and ride comfort of the vehicle.

Next, a second embodiment of the present invention will be described in detail based on the drawings. The differences between this second embodiment and the previously described first embodiment are as follows:
When calculating yaw ray 1 to correction coefficient Y, the deviation between the target yaw rate and the yaw rate is multiplied by the yaw rate in the first embodiment, whereas in the second embodiment, it is calculated by multiplying the deviation in the vehicle width direction acceleration ycc+. This is how it was configured.

Note that the system configuration and various quantities used for control are the same as in the first embodiment described above, so the same parts are denoted by the same reference numerals and the explanation will be omitted.

Next, suspension control processing, which is a feature of the second embodiment, will be explained based on the flowcharts shown in FIGS. 9(A) and 9(B). Incidentally, steps in which the same processing as the suspension control processing of the first embodiment described above is performed are indicated by the same 12-digit step number. This suspension control process is repeatedly executed at predetermined time intervals after the ECU 40 is started.

First, perform initialization processing (step 200>), read each sensor detection signal (step 210), and calculate the current heap displacement amount XHn, pitch displacement amount xpn, roll displacement!XRn, and warp displacement amount XWn at the center of gravity. 220>, heap speed 33 Once calculate DXH, pitch speed DXP, roll speed DXR, warp speed DXW Step 230), load Fl,
F2. F3. Calculate F4 Step 240 > Calculate the heap load "H" at the center of gravity, pitch torque FP, roll torque FR1 warp torque FW Step 250
>, heap target displacement amount 1-1d, pitch target displacement amount Pd at the center of gravity position.

A roll target displacement amount Rd is calculated (step 260).

Next, find the target yaw rate from the vehicle speed V and steering angle θ,
The deviation between the target yaw rate and the actually detected yaw rate γ is multiplied by the vehicle width direction acceleration Ycq to calculate the yaw rate correction coefficient Y as shown in the following equation (32), and the warp speed DXW, warp torque FW and the above A warp target displacement amount Wd at the center of gravity position is calculated from the yaw rate correction coefficient Y (step 275).

Y= (AW1xVxθ-r)XYCC]+AW2...
-(32) Next, each moon mark displacement amount Xd1. of suspension 5, 6.9.10. Xd2. Xd3. Calculate Xd4 Step 2
80>, calculate the deviations XO1, XO2, X03, XO4. Step 290>, each deviation XO1, XO2, XO
3. Output the corresponding voltage to XO4 (step 295)
. Thereafter, the process returns to step 210 above. Thereafter, in this suspension control process, steps 210 to 295 described above are repeatedly executed.

In addition, in this second embodiment, the suspension is 5°6.9
．． Reference numeral 10 corresponds to the actuator M1, and a vehicle speed sensor 27, a steering angle sensor 28, a vehicle width direction acceleration sensor 30, and a yaw rate sensor 31 correspond to the turning state detection means M2.

Further, the ECU 40 and the processes executed by the ECU 40 (steps 260, 275, 280, 290, 29
5> functions as the control means M3.

As explained above, in the second embodiment, the yaw rate correction coefficient Y is calculated by multiplying the deviation between the target yaw rate and the yaw rate γ determined from the vehicle speed V and the steering angle θ by the vehicle widthwise acceleration YCg.
is configured to calculate. Therefore, in addition to the effects of the first embodiment described above, the following effects are achieved. That is, the turning performance of the vehicle changes quickly from the basic turning characteristics to the oversteer side when steering is started, and quickly changes to the understeer side when the steering is reversed.

As shown as an example in the timing chart of FIG. 10, when steering is started at time T10, at time T11, which is slightly delayed from time T10, the vehicle width direction acceleration Ycg (indicated by a broken line in the figure) is calculated. ) occurs, and at time T12, which is further delayed, yaw rate γ (shown by a dashed line in the figure) occurs. Therefore, book 2
As in the embodiment, the yaw rate correction coefficient Y (indicated by a solid line in the figure), which is obtained by multiplying the deviation between the target yaw rate and the yaw rate γ by the vehicle width direction acceleration Ycg, starts to increase from time T11. On the other hand, as in the first embodiment described above, the yaw rate correction coefficient Y (indicated by the two-dot chain line in the figure), which is obtained by multiplying the deviation between the target yaw rate and the yaw rate γ by the yaw rate γ, is calculated from the time T11. It starts increasing from the delayed time T12.

Therefore, in the case of the second embodiment, the turning performance of the vehicle changes to the oversteer side more quickly at the start of steering than in the case of the first embodiment described above. Note that when the steering is reversed, the turning performance of the vehicle changes more quickly to the understeer side for the same reason as above. In this way book 2
According to the embodiment, the delay time required for changes in the turning performance of the vehicle at the start of steering and at the time of steering reversal can be reduced.

Furthermore, when the steering angle θ is the same, the vehicle width direction acceleration YCg increases as the vehicle speed increases. Therefore, even if the deviation between the target yaw rate and yaw rate γ is the same, the yaw rate correction coefficient Y increases as the vehicle speed increases. You can set it to a larger value depending on your needs. This has implications for improving vehicle maneuverability and stability when turning at high speeds.
It has a particularly remarkable effect.

Next, a third embodiment of the present invention will be described in detail based on the drawings. The differences between this third embodiment and the previously described first embodiment are as follows:
When calculating the yaw rate correction coefficient Y, in the first embodiment, the target yaw rate is determined from the steering angle and vehicle speed, whereas in the third embodiment, the target yaw rate is calculated based on the steering angle, vehicle speed, the square value of the vehicle speed, and vehicle characteristics. The target yaw rate is determined from a predetermined understeer setting coefficient.

Note that the system configuration and various quantities used for control are the same as in the first embodiment described above, so the same parts are denoted by the same reference numerals and the explanation will be omitted.

Next, suspension control processing, which is a feature of the third embodiment, will be explained based on the flowcharts shown in FIGS. 11(A) and 11(B). Incidentally, steps in which the same processing as the suspension control processing of the first embodiment described above is performed are indicated by the same 12-digit step number. This suspension control process is repeatedly executed at predetermined time intervals after the ECU 40 is started.

First, perform initialization processing (step 300), read each sensor detection signal (step 310), and calculate the current heap displacement amount XHn, pitch displacement amount XPn, roll displacement lXRn, and warp displacement amount XWn at the center of gravity.Step 320 ), heap speed DXH, pitch speed DXP,
Calculate roll speed DXR and warp speed DXW (Step 330), load F1. F2. F3. Calculate F4 step 340), calculate heap load FH, pitch torque FP, roll torque FR, warp torque FW at center of gravity step 350), heap target displacement amount 1-1d, pitch target displacement amount pd, roll at center of gravity position A target displacement amount Rd is calculated (step 360).

Next, the target yaw rate is determined from the vehicle speed ■, the understeer setting coefficient Kh, the vehicle speed squared value V2, and the steering angle θ, and the deviation between the target yaw rate and the actually detected yaw rate γ is multiplied by the yaw rate γ to perform yaw rate correction. The coefficient Y is calculated as shown in the following equation (33), and the warp speed DXW,
A warp target displacement IWd at the center of gravity position is calculated from the warp torque FW and the yaw rate correction coefficient Y (step 377).

Y=[(V/ (1+KhxV2)lxθXK1-7
']×γXK2+AW2...(33
) Kh...Understeer setting coefficient: 1...Actual yaw rate correction coefficient: 2...Yaw rate correction coefficient gain Here, the understeer setting coefficient Kh is a positive value determined based on vehicle characteristics, and is based on the vehicle speed. This suppresses changes in steering sensation due to increases and decreases. Further, the actually measured yaw rate correction coefficient of 1 always guarantees the mutual relationship between the vehicle speed and the steering effect (steering effect), which is determined by the above-mentioned understeer setting coefficient 1<h. A change of 1 to the measured yaw rate correction coefficient corresponds to a change in the steering gear ratio. Furthermore, the yaw rate correction coefficient dyne of 2 determines the rate of change in the front and rear wheel distribution ratio of the moving load between the left and right wheels when a deviation occurs between the target yaw rate and the yaw rate γ.

Next, each target displacement amount Xdl, Xd2. of the suspensions 5.6, 9.10. Xd3. Calculate Xd4 (step 380), deviations XO1, XO2, X03, XO4 (step 390), each deviation XOI, XO2,
Output voltage according to O3 and XO4 (step 395
). Thereafter, the process returns to step 310 above. Thereafter, in this suspension control process, steps 310 to 395 described above are repeatedly executed.

In addition, in this third embodiment, the suspension is 5°6.9
．． 10 corresponds to the actuator M1, and the vehicle speed sensor 27, the steering angle sensor 28, and the yaw rate sensor 31 correspond to the turning state detection means M2. In addition, ECU40 and the E
Processes executed by the CU 40 (steps 360, 37
7.380, 390.395> functions as the control means M3.

As explained above, the third embodiment is configured to obtain the target yaw rate from the vehicle speed V, the understeer setting coefficient 1<h, the squared vehicle speed value v2, and the steering angle θ. Therefore, in addition to the effects of the first embodiment described above, the following effects are achieved. That is, it is possible to maintain a good steering effect (steering effect) during low-speed running, and to prevent the steering effect (steering effect) from becoming excessive during high-speed running. As shown in Fig. 12, the target yaw rate (indicated by a solid line in the figure) of the third embodiment remains at a constant value as the vehicle speed V increases, depending on the value of the understeer setting coefficient 1<h. converge or decrease. Therefore, in the third embodiment, the difference in steering effect (steering effect) between low-speed travel and high-speed travel is reduced. On the other hand, as in the first embodiment described above, the target yaw rate (indicated by a broken line in the figure), which is obtained by multiplying the vehicle speed V and the steering angle θ, increases in proportion to the increase in the vehicle speed V. Therefore, in the third embodiment, for the same steering angle θ, a sufficient steering effect (steering effect) is achieved when driving at low speed, and an excessive steering effect (steering effect) is achieved when driving at high speed.
This controls the degree of steering control. In this way book 3
According to the embodiment, it is possible to improve the steering performance during high-speed driving without impairing the maneuverability of the vehicle during low-speed driving, and it is possible to guarantee the same steering feel at all times even when the vehicle speed changes.

This is particularly effective when steering the vehicle while traveling at high speeds, since it eliminates the sense of discomfort and facilitates steering.

Note that the value of the understeer setting coefficient Kh is determined based on vehicle characteristics, and for example, as shown in FIG. 12, the smaller the value of the understeer setting coefficient Kh, the more the maneuverability of the vehicle improves.

Furthermore, the larger the value of 2 is set for the yaw rate correction coefficient gain, the higher the responsiveness and followability of the control that uses the yaw rate γ as the target yaw rate.

In the third embodiment, in step 377 of the flowchart shown in FIG. 11(B), the value of the target yaw rate obtained from the vehicle speed V, the understeer setting coefficient Kh, the vehicle speed squared value V2, and the steering angle θ is used as is. Using the above equation, the yaw rate correction coefficient Y was calculated using the above-mentioned equation (33). However, for example, it is also possible to limit the calculated target yaw rate (calculated value) so that it does not exceed the allowable yaw rate (upper limit). That is, the turning performance of a vehicle is limited by the coefficient of friction between the wheels and the road surface. For example, the coefficient of friction μ between a rubber wheel and a dry asphalt pavement is approximately O99 (experimental value).
It is. The limit of the turning performance of a vehicle subject to such restrictions is approximately 1 [q] in vehicle width direction acceleration YCQ (gravitational acceleration 9.8 [m/5ec2) even if some margin is estimated.
]) degree theal.

Here, vehicle width direction acceleration YcΩ [m/5ec2], yaw rate γ [rad/sec], and vehicle speed V [m/s e
C], there is a relationship as shown in the following equation (33a).

Y CQ = 7 The turning state can be kept within the limits of the turning performance.The above-mentioned control can be realized by replacing step 377 in FIG. 11(B) with step 377A shown in FIG. 13, for example.

That is, as shown in FIG.
In a, the target yaw rate 'y'req is expressed by the following formula (33b)
Calculate as follows.

Yreq= (V/ (1+KhXV2))XθXK1
(33b) In the following step 377b, the vehicle width direction acceleration obtained by multiplying the target yaw rate Yreq calculated in the above step 377a by the vehicle speed V is added to the predetermined vehicle width direction acceleration setting constant by α (gravitational acceleration 9 It is determined whether or not the value is less than ゜8 [m/5ec2]), and if an affirmative determination is made, the process proceeds to step 377C, whereas if a negative determination is made, the process proceeds to step 3.
77d. Step 377 executed when the vehicle width direction acceleration is less than the vehicle width direction acceleration setting constant α
In C, the target yaw rate Yreq calculated in step 377a is directly set as the target yaw rate Yreq, and the process proceeds to step 377e. On the other hand, in step 377d, which is executed when the vehicle width direction acceleration is equal to or greater than the vehicle width direction acceleration setting constant α, the allowable yaw rate obtained by dividing the vehicle width direction acceleration setting constant α by the vehicle speed V is set as the target yaw rate Yreq. , and the process proceeds to step 377e.

In step 377e, the target yaw rate yreq set in step 377C or step 377d is
is used to calculate the yaw rate correction coefficient Y as shown in the following equation (33C), and the warp target displacement IWd at the center of gravity position is calculated using the yaw rate correction coefficient Y.

Y= (Yreq-7)X7XK2+AW2...(3
3c) In this way, when the vehicle width direction acceleration is limited to be less than the vehicle width direction setting constant α, the allowable yaw rate (upper limit) changes depending on changes in vehicle speed V and steering angle θ, as shown in FIG. On the other hand, it is determined as a quadratic curved surface shown by a solid line in the figure.

In addition, in the same figure, α is the vehicle width direction setting constant and the gravitational acceleration (9,
8 [m/5eC2]). On the other hand, the target yaw rate (
Calculated value) [When understeer setting coefficient Kh=0.001] is a larger value than the allowable yaw rate, as shown by the two-dot chain line in the figure. For example, the vehicle speed ■ is 60 [Km]
/h], target yaw rate (calculated value) in a turning state where the steering angle θ is 180° [Equation (33b) in step 377a
Calculated according to ]YreQ1 is 50 [deq/
sec]. By the way, this target yaw rate Yre
If the acceleration in the vehicle width direction is calculated from q1 and vehicle speed V (60 [Km/h) according to the above formula (33a>), it will be 14.54.
[m/5ec2], and the Tl width direction setting constant of 7101 degrees is equal to or more than α (9,8 [m/5ec2]). Therefore, the target yaw rate Yreq in this case is the value Yr on the curved surface that defines the allowable yaw rate shown in FIG.
eq2<approximately 34 [deg/seC]).

As mentioned above, if the vehicle width direction acceleration, which is the product of the vehicle speed V and the target yaw rate calculated from the vehicle speed V and steering angle θ, exceeds the vehicle width direction acceleration setting constant α, the vehicle width direction acceleration By configuring the setting constant to set the allowable yaw rate obtained by dividing α by the vehicle speed V as the target yaw rate Yreq, it is possible to suppress the deviation between the target yaw rate Yreq and the actually detected yaw rate γ from increasing more than necessary. . For this reason, it is possible to avoid the front wheel distribution ratio of the moving load between the left and right wheels becoming excessively small, so even when the steering angle θ is large, the so-called spin phenomenon in which the vehicle turns inside its turning radius, and the inner wheels of the turning It is possible to prevent the so-called wheel lift phenomenon, in which the rear wheels on the side lift off the road surface and become the initial stage of capsizing, and to further improve vehicle stability when turning.

Further, when the steering angle θ is small, sufficient steering effect can be ensured, and when the steering angle θ is large, stability can be improved.

Furthermore, since the front, rear, left, and right wheels are always in contact with the ground during turning, active suspension control that adjusts the distribution ratio of the load transferred between the left and right wheels to the front and rear wheels can be suitably implemented.

Incidentally, in the above case, an example was explained in which α is the gravitational acceleration of 9.8 [m/5ec2] in the vehicle width direction acceleration setting constant. However, for example, depending on the friction coefficient between the wheels and the road surface or depending on the vehicle characteristics, α in the vehicle width direction acceleration setting constant can be selected and determined from a suitable value less than or equal to the gravitational acceleration.

Next, a fourth embodiment of the present invention will be described in detail based on the drawings. The feature of the fourth embodiment is that it is configured to perform control to maintain a predetermined turning performance even if the loaded weight of the vehicle changes. That is, when detecting a turning state, the front and rear axle load sharing ratios are also determined when the vehicle is stopped or in a steady running state, and the front and rear axle load sharing ratios of the moving load between the left and right wheels are determined according to the front and rear axle load sharing ratios. control. The system configuration is the same as that of the first embodiment described above, so the same parts are denoted by the same reference numerals and the explanation will be omitted.

First, the relationship between the various quantities used for control in the fourth embodiment will be explained in the 15th embodiment.
This will be explained based on the diagram.

The various quantities detected by each of the sensors described above are the following eight quantities. That is, the displacement amount of the suspension arranged for each wheel 3.4, 7.8 x 1° x 2. X3. X4,
Load f1. f2. f3. f4 are respectively displacement converters 11,
12, 13, 14, load sensor'15, 16, 17, 1
8. Further, longitudinal acceleration XCg and vehicle width direction acceleration YCg acting on the center of gravity G of the vehicle are detected by the longitudinal acceleration sensor 29 and the vehicle width direction acceleration sensor 30. Furthermore, the vehicle speed V and steering angle θ are detected by the vehicle speed sensor 2.
7 and the steering angle sensor 28.

Based on these quantities, the motion state of the suspension arranged corresponding to each wheel 3, 4, 7.8 is calculated based on the center of gravity G of the vehicle.
into three types of motion states. In other words, heap (1-1 eav
e) Pitch, which is longitudinal vibration around an axis in the vehicle width direction passing through the center of gravity G, as indicated by an arrow P; Roll, which is rotation around an axis in the longitudinal direction indicated by an arrow R, around an axis in the longitudinal direction passing through the center of gravity G; ), there are three types of motion states.

Next, from the above three types of movement states, the deviation from the target value at the center of gravity G corresponding to each movement state is calculated. That is, there are three types: a heap vehicle height deviation HdV from a predetermined heap target vehicle height Hreq, a pitch angle deviation PdV from a pitch target angle preq, and a roll angle deviation Rdv from a roll target angle Rreq. Furthermore, the three types of deviations of the center of gravity G mentioned above are calculated as target displacements lxd'l, xd2, of each suspension provided corresponding to each wheel 3, 4, 7.8.
xd3. Convert to xd4. The ECU 40 controls each servo valve so that the displacement amount of each suspension becomes the target displacement amount. The wheelbase of the vehicle is 11. The distance between the center of gravity G and the front wheel axle is xf, the front wheel tread is Tf, and the rear wheel tread is Hr.

By the way, the front and rear wheel distribution ratio of the moving load between the left and right wheels that occurs due to the acceleration in the vehicle width direction during turning can be calculated as the roll rigidity distribution RC as shown in the following equation (34).

RC-[(Δf1-Δf2 > /((Δf1-Δf
2)+(Δf3-Δf4))]x100...(34)
However, 8f1... Left front wheel load change Δf2... Right front wheel load change Δf3... Left rear wheel load change Δf4... Right rear wheel load change Here, the steady state of vibration If we only consider this and ignore the damping term, the displacement X of each wheel and the amount of load change Δf are expressed by the following equation (35)
There is a relationship like this.

X=to/to 91. (35) However, k...spring constant Therefore, there are two methods for changing f to the amount of load change that determines the roll stiffness distribution RC, which corresponds to the front and rear wheel distribution ratio of the moving load between the left and right wheels.

That is, (1) Keep the displacement X constant and change the spring constant k.

(2) The displacement X is varied with the spring constant k kept constant.

In the fourth embodiment, the method (2) above is adopted, and the target displacement amount xd1 of each wheel. xd2. xd3. xd4, and adjust each suspension 5, 6, 9, .
10 servo valves 23, 24, 25, 26 are driven. In this way, control is performed to change the front and rear wheel distribution ratio of the load transferred between the left and right wheels during a turn to a desired value.

Next, suspension control processing executed by the ECU 40 will be explained based on FIG. 16, and roll stiffness distribution correction coefficient calculation processing will be explained based on flowcharts shown in FIG. 17.

The suspension control process shown in FIG.
After startup, it is repeatedly executed at predetermined time intervals. First, in step 400, the RAM 40c is cleared, the heap target vehicle height Hreq, pitch target angle preq, and roll target angle Rreq, which are predetermined reference values, are set, and the roll stiffness distribution correction coefficient WCOmp is initialized to the initial value O. Processing is performed. In the subsequent step 410, a process is performed in which the values obtained by A/D converting the detection signals of each of the sensors described above are read. That is, the displacement amounts x1, X2 . X
3. X4, vehicle width direction acceleration YCQ.

Eight values of vehicle speed V are read.

Next, the process advances to step 420, and as described above, the currently detected displacement amount of each suspension X1°X2. X3. X
4, the heap vehicle height H, pitch angle P, and roll angle R at the center of gravity are calculated as shown in the following equations (36) to (38).

H=X1+X2+X3+X4 ・ (36)
P=((Xl +X2)-(X3+X4))xAP'1
...(37) R= (Xi-X2)xARl + (X3-X4)XAR2...(38) However, AP
1=1/L AR1= (Xf/LAX (1/Tf>AR2= (
(L-Xf>/L) X (1/Tr) Continued step 4
In step 30, a heap vehicle height deviation 1 is calculated from the heap target vehicle height Hreq, pitch target angle Preq, and roll target angle Rreq set in step 400 and the heap vehicle height H1 pitch angle P and roll angle R calculated in step 420. −(dv1 pitch angle deviation pdV, roll angle deviation R
Processing to calculate dV is performed as shown in the following equations (39) to (41>).

Hdv=Hreq-H・・−(39) Pdv=PreQ−p・(40)
Rdv=Rreq-R-(41) Next, proceed to step 440, and calculate each deviation 1-1dv, pdv, R at the center of gravity position calculated in step 43○ above.
dV to each target displacement xd1. of the suspensions 5, 6, 9.10 arranged corresponding to each wheel 3, 4, 7.8.
xd2. xd3. A process is performed to calculate xd4 as shown in the following equations (42) to (45).

xdl-(1/4)x((Hdv+AP8xPdv)+
(AR8XRdv+WcompxYcQ))...(
42) xd2-(1/4)x((Hdv+AP8xPdv)-
(AR8xRdv+WcompxYcq))...(4
3) Xd3= (1/4)X((HdV-AP8xPdv)
+ (AR8xRdv-WcompxYcg))...
(44) xd4= (1/4)X((Hdv-AP8xPdv)
-(AR8xRdv-WcompxYcq)), ・・
・(45) ′ However, AP8=L= (1/AP1) AR8
= (LxTf)/Xf= (1/AR1>The roll stiffness distribution correction coefficient wcomp is calculated in step 400 above.
It is set to the initial value O in , and thereafter is a value calculated by the roll stiffness distribution correction coefficient calculation process described later. Also,
YCg is the acceleration in the vehicle width direction.

In the following step 450, each target displacement amount xd1. calculated in the above step 440 is calculated. xd2. After outputting a drive signal corresponding to xd3°Xd4 to each servo valve 23, 24, 25°26 of each suspension 5°6, 9, 10, the process returns to step 410.

Thereafter, this suspension control process will be performed in step 41 above.
Repeat steps 0 to 450.

Next, the roll stiffness distribution correction coefficient calculation process will be explained based on the flowchart of FIG. 17. This roll stiffness distribution correction coefficient calculation process is interrupted and executed at predetermined time intervals.

First, in step 500, it is determined whether the vehicle speed V is O or not. If the determination is affirmative, the process proceeds to step 510, and if the determination is negative, the process proceeds to step 505. In step 505, which is executed when it is determined that the vehicle is not in a stopped state, the timer t
After resetting to the value O, the mother tree roll rigidity distribution correction coefficient calculation process is ended.

On the other hand, in step 510, which is executed when it is determined in step 500 that the vehicle is in a stopped state, a time measurement process is performed in which a value of 1 is added to the value of timer t. Next step 5
15, it is determined whether or not the time value of the timer t is equal to or greater than the set time treq, and if the determination is affirmative, the process proceeds to step 525.
On the other hand, if the determination is negative, the process proceeds to step 520. In step 520, which is executed when the set time treq has not yet elapsed since the vehicle stopped, the counter n is reset to the value O, and then the process returns to step 500.

On the other hand, in step 525, which is executed when it is determined in step 515 that the vehicle has stopped for more than the set time treq after it has stopped, the load sensors 15, 16
, 17. The load f1. which is the value obtained by A/D converting the detection signal of 18. f2. A process of reading f3° f4 is performed.

Next, the process proceeds to step 530, where it is determined whether or not the counter n has been reset to the value O. If the judgment is affirmative, the process proceeds to step 535, and if the judgment is negative, the process proceeds to step 550. In step 535, which is executed when the counter n is reset to the value O, each wheel 3, 4, 7 .
8 load integrated values ΣFl, ΣF2, ΣF3. The initial value of ΣF4 is the load f1. read in step 525 above. f2
', f3. Processing to set f4 is performed. In the following step 540, a process of adding a value of 1 to the value of the counter n is performed. Next, the process proceeds to step 545, and it is determined whether the value of the counter n has reached the specified number N. If the judgment is affirmative, the process proceeds to step 555, while if the judgment is negative, the process returns to step 525. If the value of the counter n is still less than the specified number N, step 525, 53
0 and then proceeds to step 550. In step 550,
The load f1 read in step 525 above. f2. f3
．． f4 is the load integrated value ΣF1. ΣF2. ΣF3. The process of adding ΣF4 as shown in the following equations (46) to (49>) is performed.

ΣF1=ΣF1+fl...(
46) ΣF2=ΣF2+f2...
(47)ΣF3=ΣF3+f3
...(48)ΣF4=ΣF4+f4
(49) After that, the process goes through step 540 to step 545. In step 555, which is executed when the value of the counter n reaches the specified number N by repeating such load integration, the load average value FF1. of each wheel 3, 4, 7.8 is calculated. FF2. FF3. FF4 is expressed by the following formula (50)
The calculation process as shown in ~(53) is performed.

FF2=ΣF2/N...(50)
FF2=ΣF2/N...(51)
FF3-ΣF3/N...(52)
FF4-ΣF4/N...(53)
In the following step 560, the weighted average value FF1. calculated in the above step 555 is calculated. FF2. Processing is performed to calculate the axle load sharing ratio Fr of the front axle from FF3° and FF4 as shown in the following equation (54).

Fr - (FF1 + FF2) / (FFl + FF2 + FF3 + FF4) (54) Next, the process proceeds to step 565, and the roll stiffness distribution correction coefficient Wcomp is calculated from the axle load sharing ratio Fr of the front wheel axle calculated in step 560 by the following formula (55 ), then
- End the water roll stiffness distribution correction coefficient calculation process.

Wcomp=FrxK・(55)
However, after K is a coefficient determined based on the vehicle specifications, this roll stiffness distribution correction coefficient calculation process is executed by repeating steps 500 to 565 at predetermined time intervals.

In addition, in this fourth embodiment, the suspension is 5°6.9
．． 10 and the servo valves 23, 24, 25° 26 correspond to the actuator M1. In addition, load sensors 15, 1
6, 17.18 and the ECU 40 and the means executed by the ECU 40 (steps 535, 540, 545, 550,
555, 560° 565) is the turning state detection means M2, and the processing executed by the ECLI 40 and the ECU 40 (
Steps 430, 440, 450) each function as control means M3.

As explained above, in the fourth embodiment, the deviations from the target values of heap, pitch, and roll, which are three types of driving states at the center of gravity of the vehicle, are 11 dV and pdV.

Rdv is calculated and the calculated value is applied to each suspension 5°6.
9.10 target displacement xd1. xd2. xd3, xd4
and each suspension 5, according to the target displacement amount.
6.9.10, the roll stiffness distribution correction coefficient WCOmD is calculated from the axle load sharing ratio Fr of the front axle when stopped.
is calculated, and using the product of the roll stiffness distribution correction coefficient WCOmp and the vehicle width direction acceleration Ycg, the target displacement amount
d1゜xd2. xd3. It is configured to correct when calculating xd4. For this reason, when the vehicle turns, a twisting displacement of the vehicle body is generated in proportion to the acceleration in the vehicle width direction, so for example, when a single driver is on board, the front axle has a large axle load sharing ratio. When starting to turn, the rear wheel distribution ratio of the transfer load between the left and right wheels is increased to suppress the shift of the steering characteristics to the understeer side, increasing the vehicle's maneuverability. When starting a turn with a large axle load sharing ratio on the rear axle, the front wheel distribution ratio of the transfer load between the left and right wheels is increased to prevent the steering characteristics from shifting to the oversteer side, resulting in stable turning of the vehicle. guarantee. In this way, even if the load before transitioning to the turning state changes, the steering characteristics can be maintained at desired characteristics, and both the maneuverability and stability of the vehicle are improved.

In addition, as mentioned above, since the steering characteristics are controlled to suppress changes in the steering characteristics caused by fluctuations in the live load before transitioning to the turning state, there is no need to individually consider the steering characteristics under various load conditions, and the suspension characteristics By expanding the selection range, the degree of freedom when designing the suspension increases.

Furthermore, when calculating the axle load sharing ratio Fr of the front wheel axle, the load f1 of each wheel. f2. f3. f4 is integrated over a specified number of times, and the weighted average value FF1°FF2. FF3. F
The axle load sharing ratio Fr of the front axle is calculated based on F4. Therefore, it is possible to remove disturbances during load measurement due to, for example, vibrations of engine rotation, and the load average value FF1. with less error can be removed. FF2. FF3
゜Accurate front axle load sharing ratio Fr based on FF4
can be calculated, and the reliability of the calculated value is also improved.

Furthermore, after the vehicle speed once becomes zero, the load integration is started after confirming that the vehicle has been in a stopped state for a set time period treq or more.

Therefore, it is possible to prevent load accumulation in a state where the load on the front wheel axle side increases due to inertia, such as immediately after braking, and to accurately detect the load on each wheel of the vehicle in the standard state.

Note that the roll stiffness distribution correction coefficient calculation process of the fourth embodiment integrates the loads and calculates the axle load sharing ratio F of the front axle only when the vehicle is stopped continuously for a set time treq or longer.
It was configured to calculate r.

However, for example, when the vehicle is in a steady running state such as when the vehicle is running at a constant speed on a flat road, the above-mentioned loads are integrated and the axle load sharing ratio Fr of the front axle is calculated, and the roll stiffness distribution correction coefficient is calculated. It can also be configured to obtain WCOmp.

Next, a fifth embodiment of the present invention will be described in detail based on the drawings. The feature of the fifth embodiment is that it is configured to perform control to maintain sufficient stability even when the braking force is applied while the vehicle is turning. That is, when detecting the turning state, the longitudinal acceleration of the vehicle is also detected, and the front and rear wheel distribution ratio of the moving load between the left and right wheels is controlled according to the longitudinal acceleration. The system configuration is the same as in the first embodiment described above, and the various quantities used for control are the same as in the fourth embodiment, so the same parts are denoted by the same reference numerals and the explanation will be omitted.

The suspension control process, which is a feature of the fifth embodiment, will be explained based on the flowchart of FIG. 18. This suspension control process is repeatedly executed during each startup waiting period after the ECU 40 is started. First, in step 600, a process of reading vehicle specifications is performed. That is, the wheelbase L of the vehicle, the distance Xf between the center of gravity G of the vehicle and the front wheel axle,
The front wheel tread Tf and the rear wheel tread Tr are read from the ROM 40b. In the following step 610, predetermined heap target vehicle height Hreq and pitch target angle preq
, a process of reading the roll target angle Rreq is performed. Next, the process proceeds to step 620, where a process is performed in which the values obtained by A/D converting the detection signals of each of the sensors described above are read.

That is, the amount of displacement X1. X2. X3. Multi-values of X4, longitudinal acceleration XCQ, and vehicle width direction acceleration YCQ are read. In the following step 630, as described above, the currently detected displacement IX1. X2゜X3.
Based on X4, a process is performed to calculate the heap vehicle height H, pitch angle P, and roll angle R at the center of gravity as shown in the following equations (56) to (58).

H=X1+X2+X3+X4 ・ (56)
P=((X1+X2)-(X3+X4))xAPl...
・(57) R= (Xl-X2)xARl +(X3-X4) XAR2・(58) However, API=
1/L AR1= (Xf/L)x (1/Tf)AR2= (
(L-Xf>/L)x (1/Tr) Next step 64
0, and calculates the heap vehicle height deviation from the heap target vehicle height Hreq, pitch target angle preq, and roll target angle Rreq read in step 610 above and the heap vehicle height H, pitch angle P10 - roll angle R calculated in step 630 above. Hdv, pitch angle deviation pdv, roll angle deviation Rdv
is calculated as shown in the following equations (59) to (61).

Hd v-Hr eq-H・(59) Pdv=PreQ-p・(60)
Rdv=Rreq-R・ (61) In the following step 650, each deviation in the center of gravity position calculated in the above step 640) -1 dV, pdV, Rd
From V to each wheel 3, 4.7, 84. : Target displacement xd1 of correspondingly arranged suspensions 5.6, 9.10
．． xd2. xd3. Processing to calculate xd4 as shown in the following equations (62) to (65) is performed.

Xd1= (1/4)X((Hdv+AP8xPdv) +(AR
8xRdv+KKXXcaXYCQ))...(62) Xd2= (1/4)X ((Hdv+AP8XPdv)-(AR
8xRdv+KKXXcqXYC(j))...(63
) xd3= (1/4)X ((Hdv-AP8xPdv)+(AR
8XRdV-KKXXC(JXYCCI))...(6
4) Xd4= (1/4)X ((Hdv-AP8xPdv) -(A
R8xRdv-KKxXcqxYcq))...(65
) However, AP8=1= (1/AP1) AR8= (LxTf)/Xf= (1/AR1)KK
...Constant determined based on vehicle specifications Note that longitudinal acceleration XCQ is positive when decelerating, and vehicle width direction acceleration YCQ is positive when turning right.

Next, the process proceeds to step 660, and each target displacement amount xd1. xd2. After outputting drive signals corresponding to xd3 and xd4 to each servo valve 23, 24°25.26 of each suspension 5.6, 9.10, the process returns to step 620. Thereafter, in this suspension control process, steps 620 to 660 described above are repeatedly executed.

Next, an example of the above control will be explained with reference to the timing chart of FIG. 19. At time T21, vehicle 1
starts turning to the right, and vehicle width direction acceleration YCQ is generated (turning to the right is assumed to be positive). Then, the load is transferred from the right front wheel and right rear wheel, which are the inner wheels, to the left front wheel and left rear wheel, which are the outer wheels. Therefore, as shown in the figure, the left front wheel suspension load f1 and the left rear wheel suspension load f3 increase, while the right front wheel suspension load f2 and the right rear wheel suspension load f4 decrease.

Eventually, at time T22, a braking force acts on the vehicle 1, producing a longitudinal acceleration xcq (deceleration is defined as positive).

) Therefore, the load on the front wheels increases, while the load on the rear wheels decreases. In this case, if control is not performed taking into account the influence of the longitudinal acceleration It becomes easier to move into the state. However, in the fifth embodiment, when longitudinal acceleration XCCl (during deceleration) occurs, the front wheel distribution ratio of the moving load between the left and right (inner and outer) wheels is controlled to be increased. That is, by adding or subtracting the term % cg in the above formula 4, the load transferred from the right front wheel to the left front wheel is increased, while the transfer load from the right rear wheel to the left rear wheel is decreased. Therefore, each suspension load f1 to f4 is controlled as shown by the solid line in the figure, and the right rear wheel suspension load f
4 is suppressed, and transition to the locked state can be prevented. In addition, in the case of a left turn, control is similarly performed to suppress a decrease in the left rear wheel suspension load f3, which is the inner rear wheel.

In addition, in this fifth embodiment, the suspension is 5°6.9
, 10 and servo valves 23, 24, 25, 26 are 7
The longitudinal direction acceleration sensor 29 of the cutter M1 corresponds to the turning state detection means M2.

In addition, the ECU 40 and the processing executed by the ECU 40 (
Steps 640, 650, 660> function as control means M3.

As explained above, in the fifth embodiment, the deviation from the target value of the three types of motion states of heap, pitch, and roll at the center of gravity of the vehicle is −1 dV and pdv.

Hdv is calculated and the calculated value is applied to each suspension 5°6.
9.10 target displacement xd1. xd2. xd3, Xd4
and each suspension 5, according to the target displacement amount.
6, 9. When controlling 10, longitudinal direction acceleration xcq
The above target displacement amount xd'l during turning is obtained using
xd2. xd3. It is configured to calculate xd4. Therefore, when braking force is applied while the vehicle is turning,
By increasing the front wheel distribution ratio of the moving load between the left and right (outer and outer) wheels, it suppresses the decrease in the load on the inner and rear wheels, preventing the inner and rear wheels from shifting to a locked state, and improving vehicle stability during turning braking. can ensure sex.

In addition, longitudinal direction acceleration XCQ and vehicle width direction acceleration YC
Since the front and rear wheel distribution ratio of the load transferred between the left and right wheels is controlled only when the load is generated, there is no problem such as a decrease in braking performance during normal driving or a decrease in vehicle maneuverability during constant speed turning. Vehicle stability can be improved.

Furthermore, for example, in a front-wheel drive vehicle, when accelerating while turning, the rear wheel distribution ratio of the load transferred between the left and right (outer and outer) wheels increases, so the load difference between the left and right front wheels decreases, so the inner front wheel Acceleration slip can be prevented and sufficient acceleration performance can be demonstrated.

For example, in a rear-wheel drive vehicle, the equation (6
If the absolute value is used as the longitudinal direction acceleration XCQ of 2) to (65), the front wheel distribution ratio of the moving load between the left and right (inner and outer) wheels increases during acceleration during turning as well as during deceleration. Therefore, the load difference between the left and right rear wheels is reduced, so acceleration slip of the inner rear wheels during cornering can be prevented, and stable cornering acceleration can be achieved.

Although several embodiments of the present invention have been described above, the present invention is not limited to these embodiments in any way, and can be implemented in various ways without departing from the gist of the present invention. Of course.

As described in detail above, in the active suspension control device of the present invention, when the turning state detection means detects a turning state, the control means controls the front and rear wheels of the load transferred between the left and right wheels when the vehicle turns. The actuator is configured to output a signal for controlling the distribution ratio to the actuator. Therefore, the characteristics of the vehicle turning performance change depending on the actual turning condition, which has the excellent effect of improving responsiveness to steering operations during turning and increasing the stability of the vehicle attitude during turning. .

Furthermore, since the vehicle posture during turning is stabilized, the ride comfort of the vehicle is also improved.

For example, if the control means is configured to reduce the front wheel distribution ratio of the load transferred between the left and right wheels at the start of steering, and to reduce the rear wheel distribution ratio of the load transferred between the left and right wheels when the steering is reversed, The performance characteristics of the vehicle are oversteer when the steering starts, so the vehicle starts turning quickly according to the driver's intention, and understeer when the steering is reversed, allowing the vehicle to maintain a stable turning condition without excessive rotational motion. .

For example, if the control means is configured to control the front and rear wheel distribution ratio of the load transferred between the left and right wheels so that the turning state detected by the turning state detection means becomes the target turning state, The control accuracy of turning performance characteristics is improved.

Further, for example, the turning state detection means may detect steering angle, vehicle speed,
The control means detects the yaw angular velocity and the vehicle width direction acceleration, and calculates the detected vehicle width direction acceleration to the deviation between the target yaw angular velocity calculated from the detected steering angle and vehicle speed and the detected yaw angular velocity. If the front and rear wheel distribution ratio of the moving load between the left and right wheels is controlled based on the multiplied value, the turning performance of the vehicle can be quickly changed with the start of steering and the reversal of steering.

In addition, in the case of the above configuration, the amount of change in the front and rear wheel distribution ratio of the load transferred between the left and right wheels can be increased as the vehicle speed increases, which improves the maneuverability of the vehicle when turning at high speed.
It becomes possible to further improve stability and ride comfort.

For example, the turning state detecting means detects the steering angle, the vehicle speed, and the yaw angular velocity, and the control means further determines the detected steering angle, the vehicle speed, the square value of the vehicle speed, and the vehicle characteristics in advance. The vehicle is configured to control the front and rear wheel distribution ratio of the moving load between the left and right wheels based on the value obtained by multiplying the deviation between the target yaw angular velocity calculated from the understeer setting coefficient and the detected yaw angular velocity by the detected yaw angular velocity. In this case, it is possible to improve both the maneuverability of the vehicle when driving at low speeds and the steering performance when driving at high speeds, and it is possible to maintain a constant steering feeling regardless of changes in vehicle speed.

For example, when the target yaw angle calculated by the control means exceeds the allowable yaw angular velocity determined from the vehicle width direction acceleration and the vehicle speed at the time of turning, which is predetermined based on the vehicle characteristics, When the yaw angular velocity is configured to match the target yaw angular velocity, the vehicle's turning state is maintained within the limits of turning performance by suppressing excessive changes in the front and rear wheel distribution ratio of the load transferred between the left and right wheels when the vehicle turns. Therefore, it is possible to realize active suspension control that prevents the occurrence of the so-called wheel lift phenomenon and spin phenomenon and further improves stability during turning.

[Brief explanation of drawings]

FIG. 1 is a basic configuration diagram conceptually illustrating the content of the present invention, FIG. 2 is a graph showing the relationship between cornering power and load, and FIG. 3 is a system configuration diagram of the first embodiment of the present invention. Figure 4 is an explanatory diagram showing the structure of the suspension.
FIG. 5 is a block diagram illustrating the configuration of the electronic control device, FIG. 6 is an explanatory diagram illustrating the movement state of the vehicle body, and FIGS. Figure 8 is the same timing chart,
9(A) and (B) are flowcharts showing the control of the second embodiment of the present invention, FIG. 10 is a timing chart thereof, and FIGS. 11(A) and (B) are the flowcharts of the control of the second embodiment of the present invention. FIG. 12 is a flowchart showing the relationship between the target yaw rate and vehicle speed, FIG. 13 is a flowchart showing the control of a modification of the third embodiment of the present invention, and FIG. 14 is a graph showing the relationship between the vehicle speed and steering. A graph showing the relationship between the angle and the target yaw rate, FIG. 15 is an explanatory diagram showing the motion state of the vehicle body according to the fourth embodiment of the present invention, FIGS. 16 and 17 are flow charts showing the control, and FIG. A flowchart showing the control of the fifth embodiment of the present invention, and FIG. 19 is a timing chart thereof as well. Ml... Actuator M2... Turning state detection means M3... Control means 1... Vehicle 2...・Vehicle body 3.4.7.8... Wheels 5.6, 9.10... Suspension 15.16, 1
7, 18...Load sensor 27...Vehicle speed sensor 28...Steering angle sensor 29...Longitudinal acceleration sensor 30...Vehicle width direction acceleration sensor 31...Yaw rate sensor 40...Electronic control Device (ECU) 40 a-CPU

Claims (1)

[Scope of Claims] 1. An actuator disposed between each wheel of the vehicle and the vehicle body, a turning state detection means for detecting a turning state including a steering angle of the vehicle, and a turning state detected by the turning state detection means. An active suspension control device comprising: control means for outputting a signal to the actuator to control a front and rear wheel distribution ratio of a load transferred between left and right wheels of a vehicle in accordance with a turning state of the vehicle. 2. The control means according to claim 1, wherein the control means controls the front wheel distribution ratio of the load transferred between the left and right wheels to be small when steering is started, and the rear wheel distribution ratio of the load transferred between the left and right wheels is controlled to be small when the steering is reversed. Active suspension control device. 3. The active suspension control device according to claim 1, wherein the control means controls the front and rear wheel distribution ratio so that the turning state detected by the turning state detection means becomes a target turning state. 4. The turning state detection means detects a steering angle, a vehicle speed, and a yaw angular velocity, and the control means further controls a control means such that the detected yaw angular velocity becomes a target yaw angular velocity calculated from the detected steering angle and vehicle speed. The active suspension control device according to claim 1, which controls the front and rear wheel distribution ratio. 5 The turning state detection means detects a steering angle, a vehicle speed, and a yaw angular velocity, and the control means further controls a difference between a target yaw angular velocity calculated from the detected steering angle and vehicle speed and the detected yaw angular velocity. The active suspension control device according to claim 1, wherein the front and rear wheel distribution ratio is controlled based on a value obtained by multiplying the deviation by the detected yaw angular velocity. 6. The turning state detection means detects a steering angle, a vehicle speed, a yaw angular velocity, and a vehicle widthwise acceleration, and the control means detects a target yaw angular velocity calculated from the detected steering angle and vehicle speed, and a target yaw angular velocity calculated from the detected steering angle and vehicle speed. The active suspension control device according to claim 1, wherein the front and rear wheel distribution ratio is controlled based on a value obtained by multiplying the deviation from the detected yaw angular velocity by the detected vehicle width direction acceleration. 7 The turning state detection means detects a steering angle, a vehicle speed, and a yaw angular velocity, and the control means is configured to detect a steering angle, a vehicle speed, a square value of the vehicle speed, and a vehicle characteristic. Claim 1, wherein the front and rear wheel distribution ratio is controlled based on a value obtained by multiplying the deviation between the target yaw angular velocity calculated from the understeer setting coefficient and the detected yaw angular velocity by the detected yaw angular velocity. The active suspension control device described in . 8 The control means controls the allowable yaw angular velocity when the target yaw angular velocity calculated by the control means exceeds the allowable yaw angular velocity determined from the vehicle width direction acceleration and the vehicle speed at the time of turning, which are predetermined based on vehicle characteristics. The active suspension control device according to any one of claims 4 to 7, wherein the angular velocity is the target yaw angular velocity.