JP2715621B2 - Vehicle roll control device - Google Patents

Description

The present invention relates to a roll control device for a vehicle such as an automobile,
More specifically, the present invention relates to a roll control device for a vehicle including a four-wheel steering device.

[Prior Art] In a vehicle such as an automobile, for the purpose of improving small turning performance in a low vehicle speed range and improving steering stability in a middle and high vehicle speed range, Japanese Patent Application Laid-Open No. 63-17185 discloses, for example. As described, a vehicle speed-sensitive four-wheel steering device that steers the front and rear wheels so that the front and rear wheel steering angle ratio changes to a higher value in the in-phase direction than to a higher value in the opposite phase as the vehicle speed increases is already known. Have been. According to such a four-wheel steering device, the turning performance can be optimally controlled as compared with a vehicle in which only the front wheels are steered.

[Problems to be Solved by the Invention] However, in such a four-wheel steering device, since the steering angle of the rear wheels increases in the in-phase direction as the vehicle speed increases, the roll of the vehicle body is easily generated quickly, and During high-speed driving, the US-OS characteristic is set to the understeer characteristic by setting the front and rear wheel steering angle ratio to a high value in the in-phase direction,
There is a problem that the transient response of steering during high-speed running of the vehicle is poor.

The present invention has been made in view of the above-described problems in a vehicle in which the conventional four-wheel steering device is incorporated as described above, and in consideration of transient turning performance, particularly transient turning during high-speed running, while ensuring steering stability during high-speed running. An object of the present invention is to provide a roll control device for a vehicle, which is improved so that performance can be improved.

According to the present invention, there is provided a vehicle including a four-wheel steering device configured to change a front-rear wheel steering angle ratio from an opposite phase to an in-phase as the vehicle speed increases. A roll control device for front and rear wheels for controlling roll stiffness on the front wheel side and the rear wheel side, a means for determining a front and rear wheel steering angle ratio, and a means for detecting a transient turning state And control means for controlling the roll stiffness control means, wherein the control means reduces the roll stiffness on the rear wheel side as the front and rear wheel steering angle ratio increases in the in-phase direction. This is achieved by a vehicle roll control device configured to set the roll rigidity on the rear wheel side higher than that of the vehicle.

According to the configuration described above, the roll rigidity on the rear wheel side is reduced as the front and rear wheel steering angle ratio increases in the same phase direction. Is set higher.

Therefore, the US-OS characteristic can be set to the understeer characteristic without increasing the steering angle of the rear wheel when the vehicle is running at high speed, thereby preventing the vehicle body from rolling quickly and thereby improving the roll control performance of the vehicle body. Is improved. In addition, the US-OS characteristics during transient turning during high-speed running of the vehicle have a lower degree of understeer characteristics than the characteristics during steady turning, so the responsiveness during transient turning during high-speed running is improved. . In addition, US during steady turning when the vehicle is running at high speed
-OS characteristic is maintained at a predetermined understeer characteristic.
Steering stability during steady turning during high-speed running is not impaired.

The roll rigidity control means in the present invention may have any structure as long as the roll rigidity on the front wheel side and the rear wheel side can be changed independently of each other. An active suspension configured to be controllable, an active stabilizer device configured to variably control the roll rigidity, and the like may be used.

Hereinafter, the present invention will be described in detail with reference to the accompanying drawings with reference to the accompanying drawings.

Embodiment FIG. 1 is a schematic configuration diagram showing a fluid circuit of one embodiment of a roll control device according to the present invention in which a fluid pressure type active suspension is employed as roll stiffness control means. The fluid circuit of the illustrated roll control device includes actuators 1FR, 1FL, 1R provided corresponding to the right front wheel, the left front wheel, the right rear wheel, and the left rear wheel of a vehicle not shown.
R, 1RL, and these actuators have working fluid chambers 2FR, 2FL, 2RR, 2RL, respectively.

Further, in the figure, reference numeral 4 denotes a reserve tank for storing a working oil as a working fluid, and the reserve tank 4 is a suction flow passage 10 provided with a filter 8 for removing foreign matter in the middle.
Is connected to the suction side of the pump 6. The pump 6 is connected with a drain passage 12 for collecting the working fluid leaking into the pump 6 into the reserve tank 4. Pump 6
Is designed to be rotationally driven by the engine 14,
The rotation speed of the engine 14 is detected by a rotation speed sensor 16.

A high-pressure channel 18 is connected to the discharge side of the pump 6.
A check valve 20 that allows only the flow of the working fluid from the pump to each actuator is provided in the middle of the high-pressure flow path 18, and the working fluid discharged from the pump is provided between the pump 6 and the check valve 20. An attenuator 22 is provided to absorb the pressure pulsation and reduce the pressure change. One end of a front wheel high pressure passage 18F and one end of a rear wheel high pressure passage 18R are connected to the high pressure passage 18, and accumulators 24 and 26 are connected to these high pressure passages, respectively. Each of these accumulators is filled with a high-pressure gas so as to absorb the pressure pulsation of the working fluid and perform a pressure accumulating action. In the high-pressure channels 18F and 18R, the right front wheel high-pressure channel 18FR, the left front wheel high-pressure channel 18FL and the right rear wheel high-pressure channel are respectively provided.
One end of the high pressure flow path 18RL for the left rear wheel 18RR is connected.
Filters 28FR, 28FL, 28RR, 28RL are respectively provided in the middle of the high-pressure channels 18FR, 18FL, 18RR, 18RL, and the other ends of these high-pressure channels are pressure control valves 32, 34, 36, respectively.
38 pilot operated 3-port switching control valves 40, 42,
Connected to P ports 44 and 46.

The pressure control valve 32 includes a switching control valve 40, a flow path 50 that connects the high pressure flow path 18FR and the low pressure flow path 48FR for the right front wheel, and a fixed throttle 52 and a variable throttle 54 provided in the middle of the flow path. And more. The low pressure passage 48 is connected to the R port of the switching control valve 40.
FR is connected, and a connection channel 56 is connected to the A port. The switching control valve 40 is a spool valve that takes in the pressure Pp in the flow path 50 between the fixed throttle 52 and the variable throttle 54 and the pressure Pa in the connection flow path 56 as pilot pressure, and the pressure Pp is higher than the pressure Pa. When the pressure Pp and Pa are equal to each other, the switch is switched to the switch position 40b for interrupting the communication of all ports, and when the pressure Pp is lower than the pressure Pa, The switching position is switched to a switching position 40c for connecting the R and the port A to each other. The variable throttle 54 changes the effective passage cross-sectional area of the throttle by controlling the current supplied to the solenoid 58, and thereby changes the pressure Pp in cooperation with the fixed throttle 52.

Similarly, each of the pressure control valves 34 to 38 includes a pilot-operated three-port switching control valve 42, 44, 46 corresponding to the switching control valve 40 of the pressure control valve 32, and channels 60, 6 corresponding to the channel 50.
2, 64 and fixed aperture 66, 68, 70 corresponding to fixed aperture 52
And variable apertures 72, 74 and 76 corresponding to the variable aperture 54, and the variable apertures 72 to 76 are solenoids 78 and 8 respectively.
0 and 82.

The switching control valves 42, 44, and 46 are configured in the same manner as the switching control valve 40. The R ports of the switching control valves 42, 44, and 46 have a low-pressure channel 48FL for the left front wheel, a low-pressure channel 48RR for the right rear wheel, and a One end of the low-pressure flow path 48RL is connected to one end of the connection flow paths 84, 86, and 88, respectively. The switching control valves 42 to 46 are spool valves that take in the pressure Pp in the flow path 60 to 64 between the corresponding fixed throttle and the variable throttle and the pressure Pa in the corresponding connection flow path 84 to 88 as pilot pressure. When the pressure Pp is higher than the pressure Pa, the switching positions 42a, 44a,
46a, and when the pressures Pp and Pa are equal to each other, the switching positions 42b, 44b, and 46 disconnect the communication of all ports.
b, and when the pressure Pp is lower than the pressure Pa, the switching positions 42c, 44c, 46 for communicating and connecting the port R and the port A.
Switch to c.

As schematically shown in FIG. 1, each of the actuators 1FR, 1FL, 1RR, and 1RL is a cylinder 106FR, 106F, respectively.
L, 106RR, 106RL, and working fluid chambers 2FR, 2FL, 2R which are fitted to the corresponding cylinders and cooperate with the corresponding cylinders
Pistons 108FR, 108FL, 108RR, 108RL defining R, 2RL
Each is connected to a vehicle body (not shown) by a cylinder, and is connected to a suspension arm (not shown) at a tip of a rod portion of a piston. Although not shown in the drawings, a suspension spring is elastically mounted between the upper seat fixed to the rod portion of the piston and the lower seat fixed to the cylinder.

Also, cylinders 106FR, 106FL, 106R of each actuator
One ends of drain flow paths 110, 112, 114, and 116 are connected to R and 106RL. The other ends of the drain passages 110, 112, 114, and 116 are connected to a drain passage 118, and the drain passage is connected to the reserve tank 4 via a filter 120,
Thereby, the working fluid leaked from the working fluid chamber is returned to the reserve tank.

Each of the working fluid chambers 2FR, 2FL, 2RR, and 2RL has a throttle 12
Accumulators 132, 134, 13 via 4, 126, 128, 130
6, 138 are connected. Also, pistons 108FR, 108FL,
108RR and 108RL have flow paths 140FR, 140FL, 140RR, 1
40RL is provided. These flow paths connect the corresponding flow paths 56, 84 to 88 and the working fluid chambers 2FR, 2FL, 2RR, 2RL, respectively, and provide filters 142FR, 142FL, 14
It has 2RR and 142RL. Actuator 1FR, 1F
Vehicle height sensors 144FR, 144FL, 144RR, and 144RL for detecting vehicle heights XFR, XFL, XRR, and XRL at portions corresponding to the respective wheels are provided at positions close to L, 1RR, and 1RL.

Pilot-operated shut-off valves 150, 152, 154, 156 are respectively provided in the middle of the connection flow paths 56, 84 to 88, and these shut-off valves are respectively corresponding pressure control valves 40, 42, 4.
When the differential pressure between the pressure in the high pressure flow paths 18FR, 18FL, 18RR, 18RL and the pressure in the drain flow paths 110, 112, 114, 116 on the upstream side of 4, 46 is less than a predetermined value, the valve is closed. It is supposed to be maintained. Portions between the corresponding pressure control valves and the shut-off valves of the connection flow paths 56, 84 to 88 are flow paths 158, 160, respectively.
162, 164 corresponding pressure control valve flow paths 50, 60, 62,
It is connected to a portion downstream of the 64 variable throttles.
The relief valves 166, 168,
170, 172 are provided, these relief valves are respectively the upstream portion of the corresponding flow path 158, 160, 162, 164,
That is, the pressure on the side of the corresponding connection flow path is taken in as pilot pressure, and when the pilot pressure exceeds a predetermined value, the valve is opened to allow a part of the working fluid in the corresponding connection flow path to flow through the flow path 5.
0, 60-64.

The shut-off valves 150 to 156 are connected to the high pressure flow paths 18FR, 18FL, 18
The valve closing state may be maintained when the pressure difference between the pressure in RR and 18RL and the atmospheric pressure is equal to or less than a predetermined value.

The other end of the low-pressure channels 48FR and 48FL is the low-pressure channel 48F for the front wheels.
And the other ends of the low-pressure channels 48RR and RL are connected to one end of a low-pressure channel 48R for the rear wheels. The other ends of the low-pressure channels 48F and 48R are connected to one end of the low-pressure channel 48. The low-pressure channel 48 has an oil cooler 174 in the middle, and is connected at the other end to the reserve tank 4 via a filter 176. Check valve 20 and attenuator 22 in high-pressure flow path 18
Is connected to the low-pressure channel 48 through a channel 178. A relief valve 180 set at a predetermined pressure in advance is provided in the middle of the flow path 178.

In the illustrated embodiment, the high pressure passage 18R and the low pressure passage 48R
Are connected to each other by a flow path 188 having a filter 182, a throttle 184, and a normally-open type electromagnetic valve 186 capable of adjusting a flow rate. The solenoid on-off valve 186 is configured such that its solenoid 190 is excited and its exciting current is changed to open the valve and adjust the flow rate of the working fluid passing through the valve. The high-pressure passage 18R and the low-pressure passage 48R are connected to each other by a passage 194 having a pilot-operated on-off valve 192 on the way. The on-off valve 192 takes the pressure on both sides of the throttle 184 as pilot pressure, maintains the valve closing position 192a when there is no differential pressure on both sides of the throttle 184, and when the pressure on the high pressure flow path 18R side with respect to the throttle 184 is high. Valve opening position 192b
Is switched to. Thus, the throttle 184, the solenoid on-off valve 186, and the on-off valve 192 cooperate with each other to operate the high-pressure flow path 18R.
And the low-pressure flow path 48R, that is, the high-pressure flow path 18 and the low-pressure flow path 48 are selectively connected to each other to form a bypass valve 196 that controls the flow rate of the working fluid flowing from the high-pressure flow path to the low-pressure flow path.

Further, in the illustrated embodiment, the high-pressure flow path 18R and the low-pressure flow path 48R are provided with pressure sensors 197 and 198, respectively, and these pressure sensors use the pressure sensors Ps and Ps of the working fluid in the high-pressure flow path, respectively. The pressure Pd of the working fluid in the low-pressure flow path is detected. Connection channels 56, 84, 8
6 and 88 have pressure sensors 199FR, 199FL, 199RR and 19 respectively.
9RL are provided, and the pressures in the working fluid chambers 2FR, 2FL, 2RR, and 2RL are detected by these pressure sensors, respectively. Further, the reserve tank 4 has a temperature sensor for detecting the temperature T of the working fluid stored in the tank.
195 are provided.

In the vehicle to which the illustrated embodiment is applied, as shown in FIG. 17, the steering angle ratio between the front and rear wheels changes according to the vehicle speed.
A changeover switch provided in the vehicle compartment switches between a normal mode Wmn, which changes from the opposite phase to the same phase at a relatively high and low vehicle speed range, and a sports mode Wms, which changes from the opposite phase to the same phase at a relatively high vehicle speed range. A speed-sensitive four-wheel steering device configured to be operated is incorporated.

The electromagnetic on-off valve 186 and the pressure control valves 32-38 are controlled by an electric control device 200 shown in FIG. The electric control device 200 includes a microcomputer 202. The microcomputer 202 may have a general configuration as shown in FIG. 2, and includes a central processing unit (CPU) 204 and a read only memory (ROM) 2.
06, a random access memory (RAM) 208, an input port device 210, and an output port device 212, which are connected to each other by a bidirectional common bus 214.

The input port device 210 has the engine 1
4, a signal indicating the rotation speed N, a signal indicating the temperature T of the working fluid from the temperature sensor 195, a signal indicating the pressure Ps in the high-pressure passage and the signal Pd in the low-pressure passage from the pressure sensors 197 and 198, respectively. 199FL, 199FR, 199RL, 199RR from the working fluid chambers 2FL, 2FR, 2RL, 2RR, pressure Pi (i = 1,
2, 3, 4) signal, ignition switch (IG
SW) 216, a signal indicating whether the ignition switch is on or not, vehicle height sensors 144FL, 144FR, 144RL, 1
Signals indicating the vehicle heights Xi (i = 1, 2, 3, 4) of portions corresponding to the left front wheel, the right front wheel, the left rear wheel, and the right rear wheel are respectively input from 44RR.

The input port device 210 has a vehicle speed V based on the vehicle speed sensor 234.
, A signal indicating a longitudinal acceleration Ga from a longitudinal G (acceleration) sensor 236, a signal indicating a lateral acceleration Gl from a lateral G (acceleration) sensor 238, a signal indicating a steering angle θ from a steering angle sensor 240, a four-wheel steering device. A signal indicating the front-rear wheel steering angle ratio K of the four-wheel steering set by the control device 242, a vehicle height setting switch 24
Signals indicating whether the vehicle height control mode set by 8 is the high mode or the normal mode are input.

The input port device 210 appropriately processes the signal input thereto, and processes the signal based on the program stored in the ROM 206.
The processed signal is output to the CPU and the RAM 208 in accordance with the instruction of 04. ROM 206 is shown in FIG. 3, FIG. 6A to FIG.
The control flow shown in FIG. 6C and the maps shown in FIGS. 4 and 5, and FIGS. 7 to 16 are stored, and the CPU performs signal processing based on each control flow. ing.
The output port device 212 is driven by the drive circuit 220
The control signal is output to the electromagnetic switching valve 186 through the
Through 228, output control signals to the solenoids 58, 78, 80, 82 of the variable throttles 54, 72, 74, 76, respectively, and control the display 232 through the drive circuit 230. It is designed to output a signal.

Next, the operation of the illustrated embodiment will be described with reference to the flowchart shown in FIG.

Note that the control flow shown in FIG. 3 is started when the ignition switch 216 is closed. In the flowchart shown in FIG. 3, the flag Fc indicates whether the pressure Ps of the working fluid in the high-pressure passage has exceeded the threshold pressure Pc for completely opening the shut-off valves 150 to 156. 1 indicates that the pressure Ps has become equal to or higher than the pressure Pc, and the flag Fs indicates a standby pressure Pbi (i = 1, 2, 3, 4) of the pressure control valves 32 to 38 which will be described later. ) Relates to whether or not the standby pressure current Ibi (i = 1, 2, 3, 4) is set, and 1 indicates that the standby pressure current is set.

In the first step 10, a main relay (not shown) is turned on, and then in step 20.
Proceed to.

In step 20, the contents stored in the RAM 208 are cleared and all flags are reset to 0. Thereafter, the flow proceeds to step 30.

In step 30, a signal indicating the rotation speed N of the engine 14 detected by the rotation speed sensor 16, a signal indicating the temperature T of the working fluid detected by the temperature sensor 195, and a high-pressure flow detected by the pressure sensor 197. A signal indicating the pressure Ps in the passage, the pressure P in the low-pressure flow path detected by the pressure sensor 198.
Signal indicating d, pressure sensor 199FL, 199FR, 199RL, 199RR
A signal indicating the pressure Pi in the working fluid chambers 2FL, 2FR, 2RL, and 2RR detected by the signal, a signal indicating whether the ignition switch 216 is in an ON state, the vehicle height sensors 144FL, 144FR,
144RL, a signal indicating the vehicle height Xi detected by 144RR, a signal indicating the vehicle speed V detected by the vehicle speed sensor 234, the front and rear G
A signal indicating the longitudinal acceleration Ga detected by the sensor 236,
A signal indicating the lateral acceleration Gl detected by the lateral G sensor 238, a signal indicating the steering angle θ detected by the steering angle sensor 240, the front and rear wheel steering angles of the four-wheel steering set by the control device of the four-wheel steering device 242 Signal indicating ratio K, vehicle height setting switch
A signal indicating whether the mode set by 248 is the high mode or the normal mode is read, and then the process proceeds to step 40.

In step 40, it is determined whether or not the ignition switch is in an off state, and when it is determined that the ignition switch is in an off state, the process proceeds to step 200, and the ignition switch is in an on state. When the determination is made, the process proceeds to step 50.

In step 50, it is determined whether or not the engine is operating by determining whether or not the engine speed N detected by the engine speed sensor 16 and read in step 30 exceeds a predetermined value. When the determination is made and it is determined that the engine is not running, step 90 is executed.
The routine proceeds to step 60 when it is determined that the engine is operating.

The determination as to whether or not the engine is operating may be made based on whether or not the generated voltage of a generator (not shown in the drawing) driven by the engine is equal to or higher than a predetermined value.

In step 60, the operation of the timer for the time Ts from the time when the operation of the engine is started to the time when the standby pressure Pbi of the pressure control valves 32-38 is set in step 150 described below is started, Thereafter, the process proceeds to step 70. In this case, if the timer Ts has already been activated, the timer continues counting.

In step 70, the solenoid on-off valve 186 of the bypass valve 196
The current Ib supplied to the solenoid 190 is calculated in accordance with Ib = Ib + ΔIbs based on a map corresponding to the graph shown in FIG. 4 stored in the ROM 206, and then the routine proceeds to step 80.

In step 80, the current Ib calculated in step 70 is supplied to the solenoid 190 of the solenoid on-off valve 186 to drive the bypass valve 196 in the valve closing direction. Thereafter, the routine proceeds to step 90.

In step 90, the pressure Ps in the high pressure flow path is
It is determined whether or not the above is true. When it is determined that Ps ≧ Pc is not satisfied, the process proceeds to step 120, and when it is determined that Ps ≧ Pc, the process is advanced to step 100.

In step 100, the flag Fc is set to 1, and
Thereafter, the process proceeds to step 110.

In step 110, as will be described in detail later with reference to FIGS. 6A to 6C and FIGS.
The active arithmetic operation is performed based on the various signals read at the time, and the variable throttles 54, 72 to 76 of the respective pressure control valves are operated.
The current Iui to be supplied to the solenoids 58, 78, 80, 82 is calculated, and then the routine proceeds to step 170.

In step 120, it is determined whether the flag Fc is 1 or not, and it is determined that Fc = 1, that is, the pressure Ps of the working fluid in the high-pressure flow path becomes equal to or higher than the threshold pressure Pc. After that, when it is determined that the value has become lower than this, the routine proceeds to step 110, where it is determined that Fc = 1 is not satisfied, that is, the pressure Ps
When it is determined that has not exceeded the threshold pressure Pc, the routine proceeds to step 130.

In step 130, it is determined whether the flag Fs is 1 or not. When it is determined that Fs = 1, the process proceeds to step 170, and the determination that Fs = 1 is not performed. If it is, go to step 140.

In step 140, it is determined whether or not the time Ts has elapsed.If it is determined that the time Ts has not elapsed, the process proceeds to step 170, where it is determined that the time Ts has elapsed. Is performed, the routine proceeds to step 150.

In step 150, the operation of the Ts timer is stopped,
In addition, the pressure Pi read in step 30 is stored in the RAM 208 as the standby pressure Pbi, and the pressure Pi is isolated from each pressure control valve based on a map corresponding to the graph shown in FIG. Connection channels 56 and 84 between valves
The pressure of the working fluid in 88 is a standby pressure Pbi, that is, the working fluid chamber 2F detected by the corresponding pressure sensor.
In order to make the pressure substantially equal to the pressure Pi in L, 2FR, 2RL, 2RR, the variable throttles 72, 54, 7 of the pressure control valves 34, 32, 38, 36 are set.
Current Ibi applied to solenoids 78, 58, 82, 80 of 6, 74
(I = 1, 2, 3, 4) is calculated, and then the step
Proceed to 160.

In step 160, the flag Fs is set to 1 and
Thereafter, the process proceeds to step 170.

In step 170, it is determined whether or not the current Ib calculated in step 70 is equal to or more than the reference value Ibo.
When it is determined that b ≧ Ibo is not satisfied, step 30 is performed.
Returning to step 180, when it is determined that Ib ≧ Ibo, the process proceeds to step 180.

In step 180, it is determined whether or not the pressure Ps of the working fluid in the high-pressure flow path read in step 30 is equal to or higher than the reference value Pso, and it is determined that Ps is not Ps ≧ Pso. If so, the process returns to step 30, and if it is determined that Ps ≧ Pso, the process proceeds to step 190.

In step 190, the current Ibi calculated in step 150 or the current Iui calculated in step 110 is output to the solenoids 58, 78 to 82 of the variable throttles of the respective pressure control valves, so that The pressure control valve is actuated to control its control pressure, and thereafter returns to step 30 and repeats steps 30 to 190 described above.

In step 200, solenoid 190 of solenoid on-off valve 186 is used.
When the power supply to is stopped, the bypass valve 196 is opened, and thereafter, the process proceeds to step 210.

In step 210, the main relay is turned off, thereby ending the control flow shown in FIG. 3 and stopping the power supply to the electric control device 200 shown in FIG. You.

Incidentally, the pressure control by the bypass valve at the start of the above-mentioned operation is not a main part of the present invention, and the details of the pressure control are described in Japanese Patent Application No. 63-307189 filed by the same applicant as the present applicant. See issue No. The pressure control by the bypass valve when the operation is stopped may be performed as described in Japanese Patent Application No. 63-307190 filed by the same applicant as the present applicant.

Next, the active operation performed in step 110 will be described with reference to FIGS. 6A to 6C and FIGS. 7 to 16.

First, in step 300, the heave target value Rxh, the pitch target value Rxp, and the roll target value Rxr based on the target posture of the vehicle body are calculated based on maps corresponding to the graphs shown in FIGS. 7 to 9, respectively. Then, the process proceeds to step 310.

In FIG. 7, solid lines and broken lines indicate patterns when the vehicle height control mode set by the vehicle height setting switch is a normal mode and a high mode, respectively.

In step 310, the vehicle height at the position corresponding to the left front wheel, right front wheel, left rear wheel, and right rear wheel read in step 30
Based on X _{1 to} X ₄ , a displacement mode conversion operation is performed for heave (Xxh), pitch (Xxp), roll (Xxr), and warp (Xxw) in accordance with the following equations, and then the process proceeds to step 320.

_{Xxh = (X 1 + X 2} ) + (X 3 + X 4) Xxp = - (X 1 + X 2) + (X 3 + X 4) Xxr = (X 1 -X 2) + (X 3 -X 4) Xxw = (X ₁ −X ₂ ) − (X ₃ −X ₄ ) In step 320, the deviation of the displacement mode is calculated according to the following equation.

Exh = Rxh−Xxh Exp = Rxp−Xxp Exr = Rxr−Xxr Exw = Rxw−Xxw In this case, Rxw may be 0, or Xxw calculated in step 310 immediately after the start of operation of the active suspension or the past. Xxw calculated in several cycles of
May be the average value. If | Exw | ≦ W ₁ (positive constant), Exw = 0.

In step 330, the PID compensation calculation of the displacement feedback control is performed according to the following formula, and then the process proceeds to step 340.

Cxh = Kpxh · Exh + Kixh · Ixh (n) + Kdxh ｛Exh (n) −Exh (n−n ₁ )｝ Cxp = Kpxp · Exp + Kixp · Ixp (n) + Kdxp ｛Exp (n) −Exp (n−n ₁ )} Cxr = Kpxr · Exr + Kixr · Ixr (n) + Kdxr ｛Exr (n) −Exr (n−n ₁ )｝ Cxw = Kpxw · Exw + Kixw · Ixw (n) + Kdxw ｛Exw (n) −Exw (n−n ₁ )} In each of the above equations, Ej (n) (j = xh, xp, xr, xw)
Is the current Ej, Ej (n-n ₁₎ is Ej of n ₁ cycles before. Further, Ij (n) and Ij (n-1) are respectively Ij before and one cycle before, and Tx is a time constant, and Ij (n) = Ej (n) × Tx + Ij (n-1), and Ijmax is a predetermined value. | Ij | ≦ Ijmax as the value. Further, coefficients Kpj, Kij, and Kdj (j = xh, xp, xr, xw) are a proportional constant, an integral constant, and a differential constant, respectively.

In step 340, the operation of the inverse conversion of the displacement mode is performed in accordance with the following equation, and thereafter, the process proceeds to step 350.

Px ₁ = 1/4 · Kx ₁ (Cxh-Cxp + Cxr + Cxw) Px ₂ = 1/4 · Kx ₂ (Cxh-Cxp-Cxr-Cxw) Px ₃ = 1/4 · Kx ₃ (Cxh + Cxp + Cxr-Cxw) Px ₄ = 1 / 4 · Kx ₄ (Cxh + Cxp−Cxr−Cxw) where Kx ₁ , Kx ₂ , Kx ₃ and Kx ₄ are proportional constants.

In step 350, based on the maps corresponding to the graphs shown in FIG. 10 and FIG. 11, correction amounts Pga and Pgl of the pressure in the front-rear direction and the lateral direction of the vehicle are calculated, and then the step Proceed to 360.

In step 360, the pitch (Cg
p) and roll (Cpgr, Cdgr) are calculated for PD compensation of G feedforward control, and
Proceed to 370.

Cgp = Kpgp · Pga + Kdgp ｛Pga (n) −Pga (n−n ₁ )｝ Cpgr = Kpgr · Pgl Cdgr = Kdgr ｛Pgl (n) −Pgl (n−n ₁ )｝ In the above formulas, Pga (N) and Pgl (n) are the current Pga and Pgl, respectively, and Pga (nn- ₁ ) and Pgl (n
−n ₁ ) is Pga and Pgl before n ₁ cycle, respectively. Kpgp and Kpgr are proportional constants, and Kdgp and Kdgr are differential constants.

In step 370, the flow chart shown in FIG.
The steering angle read in step 30 before the cycle is θ '
The steering angular velocity is calculated according to: θ = θ−θ ′, and the change rate of the predicted lateral G based on the map corresponding to the graph shown in FIG. Is calculated, and then the process proceeds to step 380.

In step 380, on the basis of the map corresponding to the graph shown in FIG. 13, the steady-state gains Kpgrf and Kpgrr in the arithmetic expression of the operation executed in step 410 described later are calculated. Proceed to step 390.

In step 390, based on the map corresponding to the graph shown in FIG. 14, the transient gains Kdgrf and Kdgrr in the arithmetic expression of the operation executed in step 410 described later are calculated. Proceed to step 400.

In step 400, based on the map corresponding to the graph shown in FIG. 15, the gain K ₁ f of the rate of change of the estimated lateral acceleration in the arithmetic expression of the operation executed in step 410 described below. , K ₁ r are calculated, and the process then proceeds to step 410.

In step 410, the inverse G-mode conversion operation is performed according to the following equation.

In step 420, based on the pressure Pbi stored in the RAM 208 in step 150 and the result calculated in steps 340 and 410, Pui = Pxi + Pgi + Pbi (i = 1, 2, 3, 4) The target control pressure Pui of the pressure control valve is calculated,
Thereafter, the process proceeds to step 430.

In step 430, the target current to be supplied to each pressure control valve is calculated according to the following equation, and thereafter, the process proceeds to step 440.

I ₁ = Ku ₁ Pu ₁ + Kh (Psr-Ps) -Kl · Pd-α I ₂ = Ku ₂ Pu ₂ + Kh (Psr-Ps) -Kl · Pd-α I ₃ = Ku ₃ Pu ₃ + Kh (Psr-Ps −Kl · Pd I ₄ = Ku ₄ Pu ₄ + Kh (Psr−Ps) −Kl · Pd where Ku ₁ , Ku ₂ , Ku ₃ , and Ku ₄ are proportional constants for each wheel, and Kh and Kl are high pressures, respectively. A correction coefficient for the pressure in the flow path and the pressure in the low-pressure flow path, α is a correction constant between the front and rear wheels, and Psr is a reference pressure in the high-pressure flow path.

In step 440, a temperature correction coefficient Kt is calculated based on the temperature T of the working fluid read in step 30 and a map corresponding to the graph shown in FIG. 16, and Iti = Kt · Ii ( i = 1, 2, 3, 4), a temperature correction calculation of the target current is performed, and then the process proceeds to step 450.

Te is At a step _{450, Iw = (It 1 -It} 2) - (It 3 -It 4) according to the current warp (torsion amount about front-to-rear axis of the vehicle body)
Is performed, and then the process proceeds to step 460.

In step 460, the deviation of the current warp is calculated according to the following equation, using Riw as the target current warp. Thereafter, the flow proceeds to step 470.

Eiw = Riw-Iw Note that the target current warp Riw in the above equation may be zero.

In step 470, the current warp target control amount is calculated according to Eiwp = Kiwp · Eiw, using Kiwp as a proportionality constant.

In step 480, an inverse current warp conversion operation is performed in accordance with the following equation.

Iw ₁ = Eiwp / 4 Iw ₂ = −Eiwp / 4 Iw ₃ = −Eiwp / 4 Iw ₄ = Eiwp / 4 In step 490, the following equation is obtained based on the results calculated in steps 440 and 480. , The final target current Iui to be supplied to each pressure control valve is calculated, and the process then proceeds to step 170 in FIG.

Iui = Iti + Iwi (i = 1, 2, 3, 4) Thus, in the illustrated embodiment, as shown in FIGS. 13 and 14, the steady-state gain Kpgr with respect to the lateral acceleration
f, Kpgrr and transient gains Kdgrf, Kdgrr are also the front and rear wheel steering angle ratio K.
The front-wheel-side gains Kpgrf and Kdgrf are set to higher values in steps 380 and 390, respectively, as the front-rear wheel steering angle ratio K increases in the in-phase direction, and the rear-wheel-side gains Kpgrr and Kdgrr are controlled in steps 380 and 390, respectively. The front-rear wheel steering angle ratio K is set to a lower value as it increases in the in-phase direction, and each gain satisfies the following relationship.

Kpgrf> Kdgrf Kpgrr <Kdgrr Accordingly, the relationship between the front-rear wheel steering angle ratio K of the four-wheel steering device and the roll stiffness by the load distribution control of the front and rear wheels of the active suspension, that is, the US-OS characteristic is as shown in the following table. The US-OS characteristic is appropriately controlled in accordance with the control of the steering angle ratio K.

As can be seen from the table, in the in-phase region corresponding to the high speed region of the vehicle speed, the degree of understeer characteristics of the US-OS characteristic during transient turning is lower than that during steady turning, so that steering stability at high speed driving is It is possible to improve responsiveness during high-speed traveling transient turning without deteriorating the responsiveness.

In both the transitional turn and the steady turn, the US-phase is lower in the reverse phase region where the vehicle speed is lower than in the in-phase region.
Since the OS characteristics are shifted in the oversteer direction, agile turning performance is secured, and conversely, the US-OS characteristics are shifted in the understeer direction in the in-phase region where the vehicle speed is high as compared to the case of the reverse phase region. Therefore, the stability during high-speed running can be improved. Especially when the understeer of the US-OS characteristic in the reverse phase region where the vehicle speed is low is set weakly during transient turning,
Although the stability of the vehicle immediately after the turning of the steering wheel is likely to be deteriorated, in the illustrated embodiment, the US-OS characteristic at the time of steady turning in the reverse phase region is shifted in the understeer direction than during the transient turning. Therefore, it is possible to avoid deterioration of the stability of the vehicle immediately after the rotation of the steering wheel.

According to the illustrated embodiment, the rate of change of the predicted lateral acceleration calculated in step 370 is the same as the rate of change of the lateral acceleration. The gains K ₁ F and K ₁ r for the front wheel (K ₁ f) are also set to higher values as the front and rear wheel steering angle ratio increases in the in-phase direction, and thus as the vehicle speed increases, and conversely, the rear wheel side (K ₁ r) is set to be a low value, whereby the correction amount of the front wheels with respect to the rear wheels is increased as the vehicle speed becomes higher with respect to the correction of the support load based on the change rate of the predicted lateral acceleration. Accordingly, this also makes it possible to improve the stability at the time of high-speed running while securing the quick turning performance at the time of low-speed running.

Further, according to the illustrated embodiment, since each gain is controlled according to the front-rear wheel steering angle ratio K, the control mode of the four-wheel steering device is set to be smaller than when each gain is controlled according to the vehicle speed. Even when the mode is switched between the normal mode and the sports mode, the US-OS characteristics of the vehicle can be favorably controlled.

Note that the control device 200 transmits the front and rear wheel steering angle ratio K from the four-wheel steering device 242.
A signal indicating that the
200 stores a map corresponding to the graph shown in FIG. 17, and only a signal indicating whether the control mode is the normal mode or the sports mode is input from the four-wheel steering device, and the signal and the vehicle speed May be obtained by calculating the front and rear wheel steering angle ratio based on

Although the present invention has been described in detail with reference to specific embodiments, the present invention is not limited to such embodiments, and various other embodiments are possible within the scope of the present invention. That will be apparent to those skilled in the art.

[Effects of the Invention] As is clear from the above description, according to the present invention,
As the front and rear wheel steering angle ratio increases in the same phase direction, the roll stiffness on the rear wheel side is reduced, and the roll stiffness on the rear wheel side is set higher during transient turning than during steady turning.

Therefore, the US-OS characteristic can be set to the understeer characteristic without increasing the steering angle of the rear wheel during high-speed running of the vehicle, so that the roll of the vehicle body is prevented from occurring quickly,
Thereby, the roll control performance of the vehicle body can be improved. Also, the US-
Since the degree of the understeer characteristic of the OS characteristic is reduced as compared with the characteristic at the time of steady turning, the responsiveness at the time of transient turning can be improved. Further, the US-OS characteristics during steady turning during high-speed running of the vehicle are maintained at predetermined understeer characteristics, so that good steering stability during steady turning during high-speed running can be ensured.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a fluid circuit of one embodiment of a roll control device according to the present invention in which a fluid pressure type active suspension is employed as roll stiffness control means.
FIG. 3 is a block diagram showing the electric control device of the embodiment shown in FIG. 1, FIG. 3 is a flowchart showing a control flow achieved by the electric control device shown in FIG. 2, and FIG. FIG. 5 is a graph showing a map used for calculating the current Ib supplied to the bypass valve at the start of operation of the active suspension. FIG. 5 is a diagram showing the pressure Pi in the working fluid chamber of each actuator and the pressure supplied to each pressure control valve. 6A to 6C are flowcharts showing a routine of an active operation performed in step 110 of the flowchart shown in FIG. 3, and FIG. 7 is a graph showing vehicle speed V. 8 is a graph showing the relationship between the target acceleration Rxh and the target displacement Rxh, FIG. 8 is a graph showing the relationship between the longitudinal acceleration Ga and the target displacement Rxp, and FIG. 9 is a graph showing the relationship between the lateral acceleration Gl and the target displacement Rxr. Graph showing the relationship, Fig. 10 is before Graph showing the relationship between the correction amount Pga acceleration Ga and a pressure, Figure 11 is the lateral acceleration G
FIG. 12 is a graph showing the relationship between l and the pressure correction amount Pgl, and FIG. FIG. 13 is a graph showing the relationship between the front-rear wheel steering angle ratio K of four-wheel steering and steady-state gains Kpgrf and Kpgrr, and FIG. 14 is the front-rear wheel steering angle of four-wheel steering. FIG. 15 is a graph showing the relationship between the ratio K and the transient gains Kdgrf and Kdgrr. FIG. 15 shows the relationship between the front-rear wheel steering angle ratio K of four-wheel steering and the gains K ₁ f and K ₁ r of the rate of change of the estimated lateral acceleration. FIG. 16 is a graph showing the relationship between the temperature T of the working fluid and the correction coefficient Kt, and FIG.
The figure is a graph showing the relationship between the vehicle speed V and the front and rear wheel steering angle ratio K of four-wheel steering. 1FR, 1FL, 1RR, 1RL …… Actuator, 2FR, 2FL, 2R
R, 2RL… working fluid chamber, 4… reserve tank, 6… pump, 8… filter, 10… suction line, 12… drain line, 1
4… Engine, 16… Rotation sensor, 18 …… High-pressure channel, 20
…… Check valve, 22 …… Attenuator, 24,26 …… Accumulator, 32,34,36,38 …… Pressure control valve, 40,42,44,46
…… Switching control valve, 48 …… Low pressure channel, 52 …… Fixed throttle, 54
…… Variable throttle, 56 …… Connection channel, 58 …… Solenoid, 66, 6
8, 70 …… Fixed aperture, 72, 74, 76 …… Variable aperture, 78, 80, 8
2… Solenoid, 84, 86, 88 …… Connection channel, 110-118…
… Drain channel, 120… filter, 124-130… throttle, 132
~ 138 ... Accumulator, 144FR, 144FL, 144RR, 144RL
…… Vehicle height sensor, 150 to 156 …… Shutoff valve, 166 to 172 …… Relief valve, 174 …… Oil cooler, 176 …… Filter, 180 ……
Relief valve, 182… Filter, 184… Throttle, 186 …… Electromagnetic on-off valve, 190 …… Solenoid, 192 …… On-off valve, 196 …… Bypass valve, 197,198,199FR, 199FL, 199RR, 199RL …… Pressure sensor, 200 Electric control device, 202 Microcomputer, 204 CPU, ROM, 208 RAM, 210 Input port device, 212 Output port device, 216 IGSW , 220
~ 230 …… Drive circuit, 232 …… Display, 234 …… Vehicle speed sensor,
236 front and rear G sensor, 238 lateral G sensor, 240 steering angle sensor, 242 four-wheel steering device, 248 vehicle height setting switch

Claims (1)

(57) [Claims] 1. A roll control device for a vehicle including a four-wheel steering device configured to change a front-rear wheel steering angle ratio from an opposite phase to an in-phase as the vehicle speed increases. A roll stiffness control unit for front and rear wheels for controlling stiffness, a unit for calculating a front and rear wheel steering angle ratio, a unit for detecting a transient turning state, and a control unit for controlling the roll stiffness control unit. The control means reduces the rear-wheel roll stiffness as the front-rear wheel steering angle ratio increases in the same phase direction, and sets the rear-wheel roll stiffness higher during transient turning than during steady turning. Roll control device for the vehicle.