Actively controlled automotive suspension system with adjustable rolling-stability and/or pitching-stability

An actively controlled suspension system includes a vehicular rolling detecting component to adjust the suspension characteristics for suppressing the vehicular rolling. In addition, the actively controlled suspension control system may be provided for adjusting roll-stabilization load distribution between the front and rear suspension systems for adjusting vehicular driving stability by suppressing rolling and pitching of the vehicle body.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates generally to an actively controlled 
suspension system which has adjustable suspension characteristics 
according to the vehicle driving condition. More specifically, the 
invention relates to an actively controlled suspension system which has 
adjustable rolling stability and/or pitching stability. 
2. Description of the Background Art 
Generally, a typical construction of a suspension assembly comprises a 
suspension coil and shock absorber and is interposed between a vehicle 
body and a suspension member supporting a vehicular wheel, in order to 
constitute an automotive suspension system with a suspension member. The 
suspension coil spring generally resists load applied to maintain the 
vehicle body and the road wheel in a predetermined positional relationship 
to each other. On the other hand, the shock absorber is intended to damp 
or absorb vibrations transmitted between the vehicle body and the road 
wheel. The automotive suspension system may be further provided with a 
roll-stabilizer for suppressing vehicular rolling motion. 
Some of the many kinds of automotive suspension systems include hydraulic 
circuits associated with fluid chambers in hydraulic shock absorbers for 
the controlling balance between the fluid pressures in the fluid chambers 
according to the relative displacement between the vehicle body and road 
wheel. Such hydraulic circuits include a fluid pressure source supplying 
working fluid at a given pressure to the circuits, and pressure control 
valves. The pressure control valves hydraulically connect and disconnect 
the fluid pressure source to the fluid chambers of the hydraulic shock 
absorbers for controlling the pressure supply. The pressure control valves 
are controlled by an electric or electronic control system which switches 
the various valve positions to selectively introduce or drain fluid 
pressure into or from the fluid chambers so as to generate a damping force 
which suppresses vehicle body and road wheel vibrations. 
These conventional positively controlled suspension systems encounter 
various defects and have not been at all satisfactorily capable of 
suppressing vibrations or bouncing of the vehicle body in order to ensure 
riding comfort. Specifically, conventional systems produce a damping force 
by means of an orifice in the hydraulic circuit. However, due to flow 
resistance through the orifice, fluid pressure differences between the 
fluid chambers in the shock absorber cannot be easily balanced. When the 
balance is disturbed, the shock absorber tends to transmit vibration of 
the road wheel to the vehicle body which degrades riding comfort. 
In order to provide an improvement, a positively or actively controlled 
automotive suspension system was proposed in European Patent First 
Publication 01 93 124, published on Sep. 3, 1986, and assigned to the 
common owner of the present invention. The proposed positively controlled 
automotive suspension system comprises a hollow cylinder defining a 
chamber, a piston thrustingly received within the chamber of the cylinder 
and defining therein first and second fluid chambers, both filled with a 
working fluid the piston being free to move axially with the chamber, a 
fluid pressure source, a hydraulic circuit connecting the first and second 
fluid chamber and the fluid pressure source, a pressure control valve 
disposed within the hydraulic circuit and adapted to selectively establish 
and block fluid communication between the first and second fluid chambers 
and the fluid pressure source, means responsive to relative displacement 
between a vehicle body and road wheel assembly out of a predetermined 
normal range, for controlling the pressure control valve so as to adjust 
the fluid pressure in the first and second fluid chambers in order to 
adjust the relative distance between the vehicle body and the road wheel 
assembly back to within the predetermined normal range, and means 
responsive to bounding and rebounding motion of the road wheel relative to 
the vehicle body, for controlling the pressure control valve so as to 
adjust the fluid pressure in the first and second fluid chambers to assist 
smooth displacement of the piston within the cylinder thereby absorbing 
bounding and rebounding energy which would otherwise be transmitted to the 
vehicle body. 
Another type of active suspension system has been disclosed in `Autocar` 
published by Haymarket Publishing Ltd., on Sep. 10, 1987. The disclosed 
system includes a single cylinder actuator which has a cylinder tube 
connected to the vehicle body and a piston with a piston rod connected to 
the suspension member. The cylinder actuator is connected to a hydraulic 
pressure source via an electromagnetic valve. The hydraulic cylinder is 
also connected to an accumulator via an orifice. With this construction, 
the pressurized fluid to absorb road shock and suppress attitude change of 
the vehicle body is supplied from the pressure source via the pressure 
control valve. This lower response of pressure control in the hydraulic 
actuator. 
On the other hand, an automotive suspension system which has adjustable 
roll-stabilization ability was disclosed in the Japanese Patent First 
Publication (Tokkai) Showa 60-25201. The disclosed suspension system 
allows mechanical adjustment of the roll-stabilization ability by 
providing mechanical coupling in a roll-stabilizer. However, this 
adjustable roll stabilizer cannot preform precise and wide range 
roll-stability adjustment. 
SUMMARY OF THE INVENTION 
It is an object of the present invention to provide an actively controlled 
suspension system which can provide roll and/or pitching stabilization 
ability equivalent to the conventional mechanical roll-stabilizer. 
Another object of the invention is to provide an actively controlled 
suspension system which can provide wide adjustable range and precise roll 
and/or pitching-stability control. 
A further object of the invention is to provide an actively controlled 
suspension system which allows adjustment of the vehicular cornering 
characteristics by adjusting roll-stability and pitching stability 
distribution between the front and rear suspension systems. 
In order to accomplish the aforementioned and other objects, an actively 
controlled suspension system according to the present invention, includes 
a vehicular rolling detecting component to adjust the suspension 
characteristics for suppressing the vehicular rolling. In addition, the 
actively controlled suspension control system may be provided with means 
for adjusting roll-stabilization load distribution between the front and 
rear suspension systems for adjusting vehicular driving stability by 
suppressing rolling and pitching of the vehicle body. 
According to one aspect of the invention, an actively controlled suspension 
system for an automotive vehicle comprises pressure means, incorporated in 
each of suspension systems in an automotive vehicle and disposed between a 
vehicle body and a suspension member in an associated suspension system, 
for supporting the vehicle body on the suspension member, each the 
pressure means including a variable pressure chamber varying the pressure 
of a working fluid therein according to relative bounding and rebounding 
motion of vehicle body and said suspension member, a sensor means for 
monitoring kinematic acceleration exerted on the vehicle body to cause a 
change of attitude thereof, the sensor means producing an acceleration 
indicating sensor signal indicative thereof, a controller receiving the 
sensor signal to detect vehicular attitude change based thereon, and 
deriving a control signal based on the detected attitude change of the 
vehicle body for adjusting the fluid pressure in each of the variable 
pressure chambers for suppressing the attitude change of the vehicle body 
in rolling and pitching directions. 
In the preferred embodiment, the sensor means monitoring a lateral 
acceleration to be exerted on the vehicle body to produce a lateral 
acceleration indicating signal, and the controller, are responsive to the 
lateral acceleration indicating signal to perform rolling suppressive 
active suspension control for suppressing rolling motion on the vehicle 
body by adjusting fluid pressure in respective variable pressure chambers. 
The magnitudes of the increasing of the fluid pressure in one of the first 
and second variable pressure chambers substantially corresponds with the 
magnitude of decreasing of the fluid pressure in the other of the first 
and second variable pressure chambers. In the alternative, the sensor 
means comprises a vehicle speed sensor for producing a vehicle speed 
indicating signal and a steering angle sensor for producing a steering 
angle signal, and an arithmetic means for deriving a projected lateral 
acceleration based on the vehicle speed indicating signal value and the 
steering angle indicating signal value. 
In a further alternative embodiment, the sensor means monitoring a pitching 
acceleration to be exerted on the vehicle body to produce a pitching 
acceleration indicating signal and the controller, is responsive to the 
pitching acceleration indicating signal to perform pitching suppressive 
active suspension control for suppressing pitching motion on the vehicle 
body by adjusting fluid pressure in respective variable pressure chambers. 
According to another aspect of the invention, an actively controlled 
suspension system for an automotive vehicle comprises a first fluid 
pressure means provided at front portion of a first lateral side of the 
vehicle body and interposed between a vehicle body and a suspension member 
rotatably supporting a road wheel, the fluid pressure means having a first 
variable pressure chamber, a second fluid pressure means provided at the 
front portion of a second lateral side of the vehicle body and interposed 
between a vehicle body and a suspension member rotatably supporting a road 
wheel, the fluid pressure means having a first variable pressure chamber, 
a third fluid pressure means provided at rear portion of a first lateral 
side of the vehicle body and interposed between a vehicle body and a 
suspension member rotatably supporting a road wheel, the fluid pressure 
means having a first variable pressure chamber, a third fluid pressure 
means provided at rear portion or a first lateral side of the vehicle body 
and interposed between a vehicle body and a suspension member rotatably 
supporting a road wheel, the fluid pressure means having a first variable 
pressure chamber, a fourth fluid pressure means provided at the rear 
portion of a second lateral side of the vehicle body and interposed 
between a vehicle body and a suspension member rotatably supporting a road 
wheel, the fluid pressure means having a first variable pressure chamber, 
a sensor means for monitoring kinematic acceleration exerted on the 
vehicle body to cause a change of attitude thereof, the sensor means 
producing an acceleration indicating sensor signal indicative thereof, and 
a controller receiving the sensor signal to detect vehicular attitude 
change on based thereon, and deriving a control signal based on the 
detected attitude change of the vehicle body for adjusting the fluid 
pressure in each of the variable pressure chambers in the first, second, 
third and fourth pressure means for suppressing attitude change of the 
vehicle body in rolling and pitching directions. 
The sensor means may monitor a lateral acceleration to be exerted on the 
vehicle body to produce a lateral acceleration indicating signal and the 
controller is responsive to the lateral acceleration indicating signal to 
perform rolling suppressive active suspension control for suppressing 
rolling motion on the vehicle body by adjusting fluid pressure in 
respective variable pressure chambers. The sensor means may comprise a 
vehicle speed sensor for producing a vehicle speed indicating signal and a 
steering angle sensor for producing a steering angle signal, and an 
arithmetic means for deriving a projected lateral acceleration based on 
the vehicle speed indicating signal value and the steering angle 
indicating signal value. 
In the alternative embodiment, the sensor means monitoring a pitching 
acceleration to be exerted on the vehicle body to produce a pitching 
acceleration indicating signal and the controller, is responsive to the 
pitching suppressive active suspension control for suppressing pitching 
motion on the vehicle body by adjusting fluid pressure in respective 
variable pressure chambers.

DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring now to the drawing, particularly to FIG. 1, a vehicle has four 
suspension systems 11FL, 11FR, 11RL, and 11RR for respectively suspending 
vehicle body 12 on front-left, front-right, rear-left and rear-right road 
wheels 14FL, 14FR, 14RL and 14RR. Each of the front-left, front-right, 
rear-left and rear-right suspension systems 11FL, 11FR, 11RL and 11RR are 
constructed to form a so-called active suspension system which adjusts 
working fluid pressure in the suspension system for adjusting stiffness 
and damping characteristics of the suspension system in a positive manner 
for absorbing uncomfortable, relatively high frequency and relatively 
small magnitude vibration input from the road wheels 14FL, 14FR, 14RL and 
14RR and suppressing attitude change of a vehicle body. 
Each suspension system 11FL, 11FR, 11RL and 11RR comprises a suspension 
member 13, such as a suspension link, a suspension arm and so forth, and a 
suspension assembly 15 which is interposed between the vehicle body 12 and 
the suspension member 13. The suspension assembly 15 has a hydraulic 
cylinder 15A which serves as an actuator, and a coil spring 16. In the 
illustrated embodiment, the suspension coil spring 16 is not necessary to 
damp the bounding and rebounding kinematic energy and is required only to 
resiliently support the vehicle body on the road wheel resisting the 
static load due to the mass weight of the vehicle body. On the further 
positive side, the suspension coil spring 16 should be weak enough not to 
produce a damping force against vibrations to be transmitted between the 
vehicle body and the suspension member. 
The hydraulic cylinder 15A has a hollow cylinder housing 15a filled with a 
viscous working fluid and a piston 15c sealingly and thrustingly disposed 
within the internal space of the cylinder housing to divide the cylinder 
space into upper and lower fluid chambers 15d and 15e. A piston rod 15b 
extends through one end of the cylinder housing 15a. The other end of the 
cylinder housing 15a is connected to one section of the vehicle body 12. 
The lower end of the piston rod 15b is connected to the suspension member 
13. Therefore, the piston 15c is thrustingly movable in bounding and 
rebounding directions relative to the cylinder housing 15a according to 
relative displacement between the vehicle body and the suspension member. 
The hydraulic cylinder 15A of the suspension assembly 15 is connected to a 
hydraulic pressure source unit 20 via a hydraulic circuit which includes 
pressure control valve 18. The pressure control valve 18 employed in the 
shown embodiment is provided with an actuator electrically operable 
according to a suspension control signal and connected to a control unit 
30 to receive the suspension control signal therefrom. The hydraulic 
circuit includes a supply line 19s and a drain line 19d. High pressure 
accumulators 23H are connected to the supply line 19s and low pressure 
accumulators 22L are connected between the pressure control valves 18 and 
the associated hydraulic cylinder 15A. The pressure source comprises a 
pressure unit 20 and a reservoir tank 21. The pressure unit 20 is 
connected to the reservoir tank 21 to suck the viscous working fluid in 
the reservoir tank 21 to feed to the pressure control valve 18 via the 
supply line 19s. On the other hand, the drain line 19d is connected to the 
reservoir 21 to return the working fluid thereto. 
As seen from FIG. 1, the low pressure accumulators 22L are connected to a 
branch pressure line 22B connected to the communication path 27 between 
the pressure control valves 18 and the upper fluid chambers 15d of the 
hydraulic cylinder 15A. A throttle valve 22V is inserted between the 
junction of the branch pressure line 22B and the communication path 27 and 
the low pressure accumulators 22L. The throttle valve 22V has a fixed 
throttling rate to provide a predetermined flow resistance against the 
working fluid flow therethrough. 
The controller 30 is connected to a lateral acceleration sensor 29. The 
lateral acceleration sensor 29 outputs a lateral acceleration indicating 
signal .alpha.. The controller 30 derives the suspension control signals 
for controlling respective pressure control valves 18 based on the lateral 
acceleration indicating signal .alpha.. 
FIGS. 2 and 5 show the detailed construction of the hydraulic cylinder 15A 
and the pressure control valve 18. As will be seen from FIG. 2, the hollow 
cylinder housing 15a is formed with a port 15f communicating the upper 
fluid chamber 15d to an outlet port 18d of the pressure control valve 18 
via a communication line 27. The pressure control valve 18 has a valve 
housing 18A having the aforementioned outlet port 18d, an inlet port 18b 
and a drain port 18c. Respective inlet port 18b, the drain port 18c and 
the outlet port 18d are connected to a valve bore 18a defined within the 
valve housing 18A. A valve spool 19 is disposed within the valve bore 18a 
for thrusting movement therein. The valve spool 19 has first, second and 
third lands 19a, 19b and 19c. As will be seen from FIG. 2, the third land 
19c has smaller diameter than that of the first and second lands 19a and 
19b. The third land 19c defines a fifth pressure control chamber 18h which 
is connected to the drain port 18c via a drain path 18f. An actuator 
piston 22c is also disposed within the valve bore 18a. The actuator piston 
22c opposes the second land 19b in spaced apart relationship to define a 
second pressure control chamber 18i which is connected to the drain port 
18c via a drain path 18e. An annular pressure chamber 18j is defined 
between the first and second lands 19a and 19b. The pressure chamber 18j 
is constantly communicating with the outlet port 18d and thereby 
communicating with the upper fluid chamber 15d. On the other hand, the 
pressure chamber 18j shifts according to shifting of the valve spool 19 to 
selectively communicate with the inlet port 18b and the drain port 18c. A 
pressure control chamber 18k is defined between the first and third lands 
19a and 19c. The pressure control chamber 18k is in communication with the 
outlet port 18d via a pilot path 18g. A bias spring 22d is interposed 
between the actuator piston 22c and the valve spool 19. The spring force 
of the bias spring 22d balances with the hydraulic pressure in the 
pressure control chamber 18k to determine the valve spool position. The 
actuator piston 22 c contacts an actuator rod 22a of an electrically 
operable actuator 22 which comprises an electromagnetic solenoid. The 
solenoid 22 is a proportioning solenoid which varies the magnitude of 
actuation of the actuator rod 22a to determine the valve spool position. 
In order to increase the supply pressure of the working fluid, the spool 
valve 19 is shifted to the position shown in FIG. 3(A) to increase path 
area at a throttle at the inner end of the inlet port 18b by means of the 
land 19a of the spool valve 19. On the other hand, in order to decrease 
the supply pressure of the working fluid, the spool valve is shifted to 
the position shown in FIG. 3(B) to decrease the path area at the throttle 
of the inner end of the inlet port 18b and opens the drain port 18 which 
is normally blocked by means of the land 19b of the spool valve. 
Construction of the pressure control valves should not be restricted to the 
construction as illustrated in FIGS. 2, 3(A) and 3(B) but can be replaced 
with any appropriate constructions. For example, the pressure control 
valve constructions as illustrated in European Patent First Publication 01 
93 124, set forth above, can also be employed. The disclosure of the 
aforementioned European Patent First Publication 01 93 124 is herein 
incorporated by reference for the sake of disclosure. 
As seen from FIG. 2, the proportioning solenoid 22 comprises the actuator 
rod 22a and a solenoid coil 22b. The solenoid coil 22b is energized by 
suspension control signal V.sub.3 from the controller 30. The magnitude of 
energization is variable depending upon the signal level of the suspension 
control signal. Therefore, the proportioning solenoid 22 shifts the 
actuator rod in a magnitude proportional to the suspension control signal 
level. In the illustrated embodiment of the pressure control valve, the 
working fluid pressure P at the outlet port 18d is variable according to 
the characteristics shown in FIG. 4. Namely, when the suspension control 
signal V.sub.3 is zero, the pressure P at the outlet port 18 becomes 
P.sub.0 determined according to a predetermined offset pressure P.sub.0. 
When the suspension control signal value increases, the fluid pressure P 
at the outlet port 18d increases with a predetermined proportioning gain 
K.sub.1. Namely, by increasing of the suspension control valve V.sub.3, 
the actuator rod 22a is driven downwardly in FIG. 2 to the position of 
FIG. 3(A) to achieve increasing of the fluid pressure with the 
predetermined proportioning gain K.sub.1. The fluid pressure P at the 
outlet port 18d is at the output pressure P.sub.2 of the pressure unit 20. 
On the other hand, when the suspension control signal value V.sub.3 
decreases, the pressure P decreases to zero by shifting of the actuator 
rod 22a toward the direction shown in FIG. 3(B). 
The actuator rod 22a of the proportioning solenoid 22 is associated with 
the actuator with the actuator piston 22c. Contact between the actuation 
rod 22a and the actuator piston 22c can be maintained by the resilient 
force of the bias spring 22d which normally biases the actuator piston 
toward the actuation rod. On the other hand, the spring force of the bias 
spring 22d is also exerted on the valve spool 19 to constantly bias the 
valve spool downwardly in FIG. 2. The valve spool 19 also receives upward 
hydraulic force from the pressure control chamber 18k. Therefore, the 
valve spool 19 is oriented at the position in the valve bore at the 
position where the downward bias of the bias spring 22d balances with the 
upward hydraulic force of the pressure control chamber 18k. 
Here, the communication path 27, the outlet port 18d, the fluid chamber 15d 
of the hydraulic cylinder 15A and the pressure control valve 18 constitute 
the first hydraulic system. On the other hand, the low pressure 
accumulator 22L, the branch line 22B and the throttle valve 22V constitute 
the second hydraulic system with the fluid chamber 15d of the hydraulic 
cylinder 15A. The length and diameter of the pipe forming the 
communication path 27 is so selected as to generate a resistance against 
the working fluid flow therethrough. The flow resistance varies according 
to the input vibration frequency which corresponds to the stroke speed of 
the piston 15c of the hydraulic cylinder 15A in a non-linear fashion. More 
practically, the variation characteristics of the flow resistance in the 
communication path 27 is parabolic in relation to the vibration frequency. 
Furthermore, the flow resistance of the communication path is set smaller 
than the flow resistance in the second hydraulic system set forth above, 
when the input vibration frequency is lower than a broader frequency, e.g. 
7 to 8 HZ), between the resonance frequency of the vehicle body and the 
resonance frequency of the suspension member as coupled with the road 
wheel. On the other hand, the flow resistance value of the communication 
path is set greater than or equal to the flow resistance in the second 
hydraulic system set forth above, when the input vibration frequency is 
higher than or equal to the broader frequency. 
On the other hand, FIG. 6 shows the preferred embodiment of the control 
circuit for controlling the actively controlled suspension system of FIG. 
1. 
As seen from FIG. 6, the controller 30 includes control signal generator 
circuits 33f and 33r, and inverter circuits 38f and 38r. As will be seen 
from FIG. 6, the control signal generator circuit 33f is designed for 
outputting a control signal for controlling the pressure control valves 18 
in the front-left and front-right suspension systems 11FL and 11FR, which 
is, therefore, referred to as the `front suspension control signal 
generator circuit`. Similarly, the control signal generator circuit 33r is 
designed for outputting the control signal for controlling the pressure 
control valves 18 in the rear-left and rear-right suspension systems 11RL 
and 11RR, which will be hereafter referred to as `rear suspension control 
signal`. 
The control signal generator circuits 33f and 33r comprise gain-controlled 
amplifiers respectively having variable gains Kf and Kr. On the other 
hand, the control signal generator circuits 31f and 31r are connected to a 
roll-stability distribution derivation circuit 34. The gain control signal 
generator circuit 34 is, in turn, connected to a total roll-stability 
setting circuit 35, and a front/rear distribution setting circuit 36. The 
front/rear distribution setting circuit 36 is connected to a steering 
angle sensor 37 and a manual selector unit 36A which includes a mode 
selector switch 36a and a manual set switch 36b. In the preferred 
construction, the total roll-stability setting circuit 35 includes a 
manual selector switch provided adjacent the driver's seat in the vehicle 
cabin so that the driver may be able to reach for setting the desired 
overall stiffness of the suspension systems and thereby setting the total 
roll-stability. The total roll-stability setting circuit 35 outputs a set 
total roll-stability indicating signal Z.sub.0 to the gain control signal 
generator circuit 34. 
On the other hand, the front/rear distribution setting circuit 36 receives 
a mode selector signal form the manual selector unit 36A. In practice, the 
mode selector switch 36a of the manual selector unit 36A is operable for 
manually selecting one of an AUTO mode and MANUAL mode. The manual set 
switch is operable when the mode selector switch 36a is set at the MANUAL 
mode position for setting front/rear distribution of the roll-stability. 
In the AUTO mode as selected by means of the mode selector switch 36a, the 
front-rear distribution setting circuit 36 is active to automatically 
adjust the front/rear distribution of the roll-stability in accordance 
with a predetermined front/rear distribution adjusting parameter. In the 
shown embodiment, the steering angle is taken as the front/rear 
distribution adjusting parameter. Therefore, the front/rear distribution 
setting circuit 36 is connected to the steering angle sensor 37 to receive 
therefrom a steering angle sensor. The steering angle sensor 37 may have 
per se well known construction and may comprise a photoelectric sensor, a 
potentiometer and so forth. 
It should be appreciated that though the steering angle sensor 37 is 
employed in the illustrated embodiment as a sensor for detecting vehicular 
rolling condition and as the front/rear distribution adjusting parameter 
detecting means, it would be possible to employ other sensors which can 
satisfactorily detect vehicular rolling condition, such as a lateral 
acceleration sensor, a lateral force sensor and so forth. Furthermore, the 
manual selector unit 36A may be provided adjacent the drives's seat so 
that the driver may easily reach the same. 
The front/rear distribution setting circuit 36 derives front/rear 
distribution based on the inputs from the manual selector unit 36A and the 
steering angle sensor 37. The front/rear distribution setting circuit 36 
outputs a front/rear distribution indicating signal .alpha. to the gain 
control signal generator circuit 34. 
The gain control signal generator circuit 34 is responsive to the total 
roll-stability indicating signal Z.sub.0 and the front/rear distribution 
indicating signal .alpha. to derive gain control signals Zf and Zr. The 
gain control signal Zf to be transmitted to the front suspension control 
signal generator circuit 33f for determining the amplifier gain Kf, which 
gain control signal for adjusting the amplifier gain Kf in the front 
suspension control signal generator circuit will be hereafter referred to 
as the `front gain control signal`. On the other hand, the gain control 
signal Zr is transmitted to the rear suspension control signal generator 
circuit 33r for determining the amplifier gain Kr, which gain control 
signal will be hereafter referred to as the `rear gain control signal`. In 
practical operation, the front/rear distribution setting circuit 36 
processes the manually set value indicating signal from the manual set 
switch 36b in the manual selector unit 36A as the mode selector switch 36a 
is set in the MANUAL mode position, the mode selector signal from the mode 
selector switch and the steering angle signal from the steering angle 
sensor 37. When the MANUAL mode is selected through the mode selector 
switch 36a, the manually set value as represented by the manually set 
value indicating signal is output from the front/rear distribution setting 
circuit 36 as the front/rear distribution indicating signal .alpha.. On 
the other hand, when the AUTO mode is selected, the front/rear 
distribution setting circuit 36 derives the front/rear distribution value 
.alpha. according to the vehicular steering condition, automatically. In 
practice, the front/rear distribution setting circuit 36 sets the 
front/rear distribution value .alpha. at a predetermined initial value 
which is greater than or equal to 0.5, while the vehicle travels straight. 
On the other hand, the front/rear distribution setting circuit 36 is 
responsive to initiation of the vehicular steering operation which is 
detected by monitoring a change of the steering angle signal, to set the 
front/rear distribution value .alpha. at a value in the range of 0 to 0.5. 
The front/rear distribution setting circuit 36 is also responsive to 
termination of the vehicular steering operating which is detected by 
monitoring a change of the steering angle signal, to set the front/rear 
distribution value .alpha. at a value in the range of 0 to 0.5. The 
relationship of the front and rear gain control signal values Zf and Zr 
with respect to the front/rear distribution indicating value .alpha. may 
be seen from FIG. 6. 
The gain control signal generator circuit 34 receives the total 
roll-stability indicating signal Z.sub.0 from the total roll-stability 
setting circuit 35 and the front/rear distribution indicating signal 
.alpha. from the front/rear distribution setting circuit 36. The gain 
control signal generator circuit 34 processes the inputs, i.e. Z.sub.0 and 
.alpha. to derive the front gain control signal Zf and the rear gain 
control signal Zr according to the following equations: 
EQU Zf=.alpha..times.Z.sub.0 (1) 
EQU Zr=Z.sub.0 -Zf (2) 
The front suspension control signal generator circuit 33f receives the 
lateral acceleration indicating signal .alpha. from the lateral 
acceleration sensor 29. The front suspension control signal generator 
circuit 33f also receives the front gain control signal Zf from the gain 
control signal generator circuit 34. The front suspension control signal 
generator circuit 33f determines the amplifier gain Kf according to the 
front gain control signal value Zf and amplifies the lateral acceleration 
indicating signal value .alpha. the determined gain Kf in order to derive 
the front suspension control signal Cf. The front suspension control 
signal Cf is fed to the proportioning solenoid 22 of the pressure control 
valve 18 in the front-right suspension system 11FR. The front suspension 
control signal Cf is also fed to the proportioning solenoid 22 of the 
pressure control valve in the front-left suspension system 11FL via the 
inverter circuit 38f. Therefore, the different polarity of the front 
suspension control signals Cf are input to the solenoids 22 of the 
pressure control valves 18 in the front-right and front-left suspension 
systems 11FR and 11FL. 
The rear suspension control signal generator circuit 33r receives the 
lateral acceleration indicating signal .alpha. from the lateral 
acceleration sensor 29. The rear suspension control signal generator 
circuit 33r also receives the rear gain control signal Zr from the gain 
control signal generator circuit 34. The rear suspension control signal 
generator circuit 33r determines the amplifier gain Kr according to the 
rear gain control signal value Zr and amplifies the lateral acceleration 
indicating signal .alpha. with the determined gain Kr in order to derive 
the rear suspension control signal Cr. The rear suspension control signal 
Cr is fed to the proportioning solenoid 22 of the pressure control valve 
18 in the rear-right suspension system 11RR. The rear suspension control 
signal Cr is also fed to the proportioning solenoid 22 of the pressure 
control valve in the rear-left suspension system 11RL via the inverter 
circuit 38f. Therefore, the different polarity of the rear suspension 
control signals Cr are input to the solenoids 22 of the pressure control 
valves 18 in the rear-right and rear-left suspension systems 11RR and 
11RL. 
Assuming the load distribution at respective road wheels 14FL, 14FR, 14RL 
and 14RR are even, characteristics of respective hydraulic cylinders 15A, 
of the loop-gains of the hydraulic pressure control circuits and of coil 
springs are same, distribution of the rolling stability at the front and 
rear suspension systems can be controlled by adjusting the fluid pressures 
in the hydraulic cylinders according to the front and rear suspension 
control signals Cf and Cr. By controlling the roll-stability distribution 
at the front and rear suspension systems, vehicular steering 
characteristics can be adjusted. 
Namely, when the front suspension control signal Cf has greater value than 
that of the rear suspension control signal Cr, the roll-stabilization load 
at the front suspension systems 11FL and 11FR becomes greater than that of 
the rear suspension systems 11RL and 11RR. This causes reducing the 
relative cornering force as a total cornering force of the front-left and 
front-right suspension systems 11FL and 11FR at the front suspension 
systems to be smaller than that in the rear suspension systems. This 
increases cornering factor Ks to increase understeering characteristics of 
the vehicle. On the other hand, when the rear suspension control signal Cr 
has greater value than that of the front suspension control signal Cf, the 
roll-stabilization load at the rear suspension systems 11RL and 11RR 
becomes greater than that of the front suspension systems 11FL and 11FR. 
This causes reducing the relative cornering force as a total cornering 
force of the rear-left and rear-right suspension systems 11RL and 11RR at 
the rear suspension systems to be smaller than that in the front 
suspension systems. This decreases cornering factor Ks to increase 
oversteering characteristics of the vehicle. 
Operation of the preferred embodiment of the actively controlled suspension 
system according to the invention will be described here below. 
At first, it will be appreciated when the front suspension control signal 
value Cf is substantially same as rear suspension control signal value Cr, 
the roll-stabilization load distribution at front and rear suspension 
systems becomes even to provide substantially neutral steering 
characteristics. 
As will be appreciated, in general, the resonance frequency of the 
suspension member with the road wheel is higher than the resonance 
frequency of the vehicle body. Therefore, when the vibration is input from 
the suspension member, the vibration frequency is usually higher than the 
broader frequency. On the other hand, when the vehicle body causes 
rolling, pitching, bouncing or so forth to input vibration, the vibration 
frequency is lower than the broader frequency. Since the flow resistance 
of the communication path 27 becomes greater than that of the second 
hydraulic system when the vibration is input from the suspension member. 
Therefore, in this case, the second hydraulic system becomes active to 
absorb vibration energy. On the other hand, when the vibration frequency 
is lower than the broader frequency during vehicular attitude change, such 
as vehicular rolling, pitching and bouncing and so forth, the first 
hydraulic system is active to adjust the fluid pressure in the fluid 
chamber 15d to suppress attitude change of the vehicle body. 
For example, when bounding motion occurs at the suspension member, the 
piston 15c of the hydraulic cylinder 15A shifts upwardly to cause 
increasing of the fluid pressure in the upper chamber 15d. Since the input 
vibration frequency is higher than the broader frequency, the increased 
pressure is introduced into the low pressure accumulator 22L through the 
throttle valve 22V because that the second hydraulic circuit has lower 
flow resistance than that of the first hydraulic circuit. In this case, 
the throttle valve generates damping force against the piston stroke to 
successfully prevent the vibration energy input from the suspension member 
from being transmitted to the vehicle body. 
Therefore, in response to the bounding motion of the suspension member 
causing increasing of the fluid pressure in the fluid chamber 15d, the 
pressurized fluid flows from the fluid chamber 15d to the low pressure 
accumulator 22L via the branch line 22B and the throttle valve 22V. Since 
the throttle valve 22V has a given throttling rate to limit fluid flow 
therethrough, this flow resistance serves as damping force for absorbing 
vibration energy so that the vibration energy is not transmitted to the 
vehicle body. 
While the vehicle travels steadily as set forth above and assuming the AUTO 
mode is selected through the mode selector switch 36a of the manual 
selector unit 36A, the stroke indicating signals from respective stroke 
sensors 29FL, 29FR, 29RL and 29RR are substantially the same. Therefore, 
the outputs of the subtractor circuits 31 and 32 are held substantially 
zero. As a result, the front and rear suspension control signals Cf and Cr 
become substantially zero. Therefore, the solenoid coils 22a of the 
proportioning solenoids 22 of the pressure control valves 18 are held 
deenergized to maintain the fluid pressure in the fluid chambers 15d of 
respective hydraulic cylinders 15A at the initial offset pressure P.sub.0. 
At this condition, the aforementioned first hydraulic systems in 
respective suspension systems are active for absorbing the road shock and 
other relatively high frequency vibrations. 
In the alternative, it would be possible to avoid the first hydraulic 
system and absorb the road shock and so forth by the action of the valve 
spool 19 in response to pressure variation in the pressure control valve 
18K. Namely in this case, when bounding motion occurs at the suspension 
member, the piston 15c of the hydraulic cylinder 15A shifts upwardly to 
cause increasing of the fluid pressure in the upper chamber 15d. This 
causes increasing of the fluid pressure at the outlet port 18d of the 
pressure control valve 18. As a result, the fluid pressure in the pressure 
control chamber 18k increases by the pressure introduced through the pilot 
path 18g to destroy the balance between the downward bias of the bias 
spring 22d and the upward hydraulic force of the pressure control chamber 
18k. This causes upward movement of the valve spool 19 against the spring 
force of the bias spring 22d, as shown in FIG. 3(B). As a result, path 
area of the drain port 18c increases and the inlet port 18b becomes being 
blocked. Therefore, the fluid pressure in the fluid chamber 15d is drained 
through the drain port. Therefore, the increased fluid pressure in the 
fluid chamber 15d of the hydraulic cylinder 15A can be successfully 
absorbed so that the bounding energy input from the suspension member will 
not be transmitted to the vehicle body. On the other hand, when rebounding 
motion occurs at the suspension member, the piston 15c of the hydraulic 
cylinder 15A shifts downwardly to cause decreasing of the fluid pressure 
in the upper chamber 15d. This causes decreasing of the fluid pressure at 
the outlet port 18d of the pressure control valve 18. As a result, the 
fluid pressure in the pressure control chamber 18k decreases by the 
pressure introduced through the pilot path 18g to destroy the balance 
between the downward bias of the bias spring 22d and the upward hydraulic 
force of the pressure control chamber 18k. This causes downward movement 
of the valve spool 19 against the spring force of the bias spring 22d, as 
shown in FIG. 3(A). As a result, path area of the inlet port 18b increases 
and the drain port 18c becomes being blocked. Therefore, the fluid 
pressure in the fluid chamber 15d is increased by the pressure introduced 
through the inlet port. Therefore, the decreased fluid pressure in the 
fluid chamber 15d of the hydraulic cylinder 15A can be successfully 
absorbed so that the rebounding energy input from the suspension member 
will not be transmitted to the vehicle body. 
If steering operation is performed for making left-hand turn, the vehicle 
body rolls to lower the left side and to lift-up the right side due to 
lateral force exerted on the vehicle body because of the presence of the 
centrifugal force so that the vehicle body inclines toward left with an 
angle .theta. with respect to the vertical plane, as shown in FIGS. 5, 7 
and 8. Therefore, the lateral acceleration sensor 29 outputs the lateral 
acceleration indicating signal .alpha.. If the front suspension systems 
are concerned, the fluid pressures in the hydraulic cylinders 15A of the 
front-left and front-right suspension systems 11FL and 11FR are determined 
according to the following process. 
Assuming the mass weight of the vehicle body is M, the effective area of 
the hydraulic cylinder 45A of the front-left and front-right suspension 
systems is A, and the fluid pressure variation gain in the pressure 
control valve 18 is K.sub.1, the variation magnitude of the fluid pressure 
in the hydraulic cylinder in response to the front suspension control 
signal Cf may be illustrated by the following equation (3): 
EQU P-P.sub.0 =K.sub.1 .times.Cf (3) 
Here, setting the (P-P.sub.0) as .DELTA.P, the foregoing equation can be 
rewritten as the following equation (4): 
EQU .DELTA.P=K.sub.1 .times.Cf (4) 
Since the hydraulic force to be generated in each cylinder is A (effective 
area).times..DELTA.P, the rolling moment Mr can be illustrated by the 
following equation: 
EQU Mr=A.times..DELTA.P+K(x.sub.2 -x.sub.1) (5) 
where 
x.sub.2 is the vehicle body elevation; 
x.sub.1 is the suspension member elevation; and 
K is a spring coefficient of the coil spring 16. 
Since the front suspension control signal value Cf is derived by 
(Kf.times..alpha.), the foregoing equation (5) can be modified as: 
EQU Mr+K(x.sub.2 -x.sub.1)=K.sub.1 .times.Cf.times.A (6) 
Assuming the magnitude x.sub.1 of the displacement of the suspension member 
is zero, the magnitude x.sub.2 of the displacement of the vehicle body can 
be illustrated in a form of vibration transmission coefficient by the 
following equation (7) with Laplace conversion: 
EQU x.sub.2 /Cf=(K.sub.1 .times.A)/(MS.sup.2 +K) (7) 
Therefore, the vibration transmission coefficient (x.sub.2 /.alpha.) can be 
illustrated by the following equation (8): 
EQU x.sub.2 /.alpha.=(K.sub.1 .times.Ax Kf)/(MS.sup.2 +K) (8) 
On the other hand, FIG. 9 shows the vehicle to which anti-roll suspension 
control is not effected. In this case, the rolling motion of the vehicle 
body can be illustrated by the following equation (9): 
EQU J.theta.+K.times.(L.sup.2 /2).times..theta.=M.times.H.times..alpha.(9) 
where 
J is a rolling inertia moment; 
.theta. is rolling angle; 
H is a distance between the gravity center and roll center; 
K is a spring coefficient; 
L is a tread distance. 
The foregoing equation (9) can be modified as: 
EQU .theta./.alpha.=(M.times.H)/(JS.sup.2 +KL.sup.2 /2) (10) 
Since x.sub.2 is (L.theta./2), the foregoing equation (10) can be further 
modified as: 
EQU X.sub.2 /.alpha.=(L.times.H.times.M/2)/(JS.sup.2 +KL.sup.2 /2)(11) 
As will be appreciated the foregoing equation (8) illustrates the response 
characteristics of the hydraulic system of the shown embodiment and the 
equation (11) illustrates the rolling motion of the vehicle body in 
response to the lateral acceleration. Therefore, as will be seen from the 
equations (8) and (11), the rolling motion of the vehicle body can be 
dynamically suppressed by setting the amplifier gain kf appropriately. 
This can be regarded as substantially equivalent to the roll-stabilizing 
force to be created by the conventional roll-stabilizers. 
On the other hand, upon initiation of the steering operation, the 
front/rear distribution setting circuit 36 detects the fact based on 
variation of the steering angle signal from the steering angle sensor 37. 
As set forth, since the AUTO mode is selected, the front/rear distribution 
setting circuit 36 sets the front/rear distribution indicating value 
.alpha. within a range of 0 to 0.5 for a given period of time. Therefore, 
front gain control signal Zf to be fed to the front suspension control 
signal generator circuit 33f becomes smaller than the rear suspension 
control signal Zr which is to be fed to the rear suspension control signal 
generator circuit 33r. Therefore, the front amplifier gain Kf of the front 
suspension control signal generator circuit 33f becomes smaller than the 
rear amplifier gain Kr of the rear suspension control signal generator 
33r. Assuming the rolling magnitude at front suspension systems 11FL and 
11FR and the rear suspension systems 11RL and 11RR are the same, the fluid 
pressure variation .DELTA.P in the front suspension systems becomes 
smaller than that of the rear suspension systems. This increases 
oversteering characteristics of the vehicle to provide better turning 
ability. 
In the practical suspension control for the front suspension systems 11FL 
and 11FR, the front suspension control signals Cf which becomes the 
negative value, is fed to the proportioning solenoid 22 of the pressure 
control valve 18 in the front-left suspension system 11FL via the inverter 
38f. Therefore, the proportioning solenoid 22 becomes active to operate 
the pressure control valve 18 in the front-left suspension system to 
increase the fluid pressure in the front-left suspension system to raise 
the left side of the vehicle body toward the neutral position. On the 
other hand, the negative value of front suspension control signals Cf 
negative value is directly fed to the proportioning solenoid 22 of the 
pressure control valve 18 in the front-right suspension system 11FR. 
Therefore, the proportioning solenoid 2 becomes active to operate the 
pressure control valve 18 in the front-right suspension system to decrease 
the fluid pressure in the front-right suspension system to lower the right 
side of the vehicle body toward the neutral position. Therefore, the 
vehicle rolling can be successfully suppressed. 
Similarly, the rear suspension control signal Cr which is derived 
substantially in the same manner as that for the front suspension control 
signal and has greater absolute value than that of the front suspension 
control signal. The rear suspension control signals Cr which becomes the 
negative value, is fed to the proportioning solenoid 22 of the pressure 
control valve 18 in the rear-left suspension systems 11RL via the inverter 
38r. Therefore, the proportioning solenoid 22 becomes active to operate 
the pressure control valve 18 in the rear-left suspension system to 
increase the fluid pressure in the rear-left suspension system to raise 
the left side of the vehicle body toward the neutral position. On the 
other hand, the negative value of rear suspension control signals Cr is 
directly fed to the proportioning solenoid 22 of the pressure control 
valve 18 in the rear-right suspension system 11RR. Therefore, the 
proportioning solenoid 22 becomes active to operate the pressure control 
valve 18 in the rear-right suspension system to decrease the fluid 
pressure in the rear-right suspension system to lower the right side of 
the vehicle body toward the neutral position. Therefore, the vehicle 
rolling can be successfully suppressed. 
On the other hand, upon termination of the steering operation, the 
front/rear distribution setting circuit 36 detects the fact based on 
variation of the steering angle signal from the steering angle sensor 37. 
As set forth, since the AUTO mode is selected, the front/rear distribution 
setting circuit 36 sets the front/rear distribution indicating value 
.alpha. within a range of 0.5 to 1 for a given period to time. Therefore, 
the front gain control signal Zf to be fed to the front suspension control 
signal generator circuit 33f becomes greater than the rear suspension 
control signal Zr which is to be fed to the rear suspension control signal 
generator circuit 33r. Therefore, the front amplifier gain Kf of the front 
suspension control signal generator circuit 33f becomes greater than the 
rear amplifier gain Kr of the rear suspension control signal generator 
33r. Assuming the rolling magnitude at front suspension systems 11FL and 
11FR and the rear suspension systems 11RL and 11RR are the same, the fluid 
pressure variation .DELTA. in the front suspension systems becomes smaller 
than that of the rear suspension systems. This increases understeering 
characteristics of the vehicle to provide better stability. 
Therefore, according to the shown embodiment, the active suspension system 
according to the present invention, exhibits roll-stabilization ability 
equivalent to the conventional mechanical roll-stabilizer. In addition to 
this, the shown embodiment of the active suspension system can provide 
improved cornering characteristics with variation of the steering 
characteristics during cornering. 
In addition, the shown embodiment is provided capability of manually 
setting the front/read distribution value in MANUAL mode, steering 
characteristics fitting the driver's taste can be selected. 
FIG. 11 shows a modified embodiment of the actively controlled suspension 
system according to the invention. In this embodiment, the lateral 
acceleration sensor 29 is replaced with a lateral acceleration projecting 
means 40. This is intended to avoid influence of the lateral acceleration 
generated due to vehicular rolling to perform anti-rolling active 
suspension control purely depending on the lateral acceleration externally 
exerted on the vehicle body due to the centrifugal force generated by 
cornering. In addition, by utilizing the lateral acceleration projecting 
means 40, self-excited vibration which otherwise occurs in the lateral 
acceleration sensor when the sensitivity is increased, can be avoided. 
In the shown embodiment, the lateral acceleration projecting means 40 
comprises a vehicle speed sensor 41, a steering condition detecting means 
42 which includes a steering angle sensor 42a and an actual steering angle 
derivation circuit 42b, and an arithmetic circuit 43. The vehicle speed 
sensor 41 monitors the vehicle traveling speed to produce a vehicle speed 
indicating signal V. In practice, the vehicle speed sensor 41 is designed 
to monitor rotation speed of a power transmission output shaft. A 
transmission output shaft reference position indicating pulse is produced 
at every predetermined angular position of the transmission output shaft. 
The transmission output shaft reference position indicating pulse is 
counted to derive the vehicle speed value based on the counter value in 
the given unit time. The steering angle sensor 42a monitors steering 
angular displacement to produce a steering angle signal .delta..sub.0. The 
steering angle sensor 42a is connected to the actual steering angle 
derivation circuit 42b to feed the steering angle signal .delta..sub.0. 
The actual steering angle derivation circuit 42b derives the actual 
steering angle .delta. based on the steering angle signal value and a 
known steering gear ratio N. In practice, the actual steering angle 
.delta. can be derived from the following equation: 
EQU .delta.=.delta..sub.0 /N 
Based on the above, the arithmetic circuit 43 performs calculation of the 
actual lateral acceleration G according to the following equation (12): 
##EQU1## 
C.sub.1 is a cornering force at the front wheel; C.sub.1 is a cornering 
force at the rear wheel; 
l is wheel base length; 
l.sub.1 is a distance between the front wheels and a gravity center; 
l.sub.2 is a distance between the rear wheels and a gravity center; 
M is a mass weight of the vehicle; 
I is yawing moment to be exerted on the vehicle; 
K.sub.s is a stability factor; 
S is a Laplace conversion coefficient; 
.zeta..sub.1, .zeta..sub.2 are damping ratios; and 
.omega..sub.1 and .omega..sub.2 are natural frequencies. 
According to the foregoing equation (12), the relationship between the 
logarithm of the G/.delta. and the vehicle V can be illustrated as shown 
by the solid line L.sub.1 in FIG. 12. As will be seen herefrom, the 
variation characteristics of the solid line L.sub.1 is equivalent to that 
of the broken line L.sub.2 which shows actually measured lateral force on 
the vehicle. On the other hand, the relationship between the frequency of 
the bounding and rebounding motion and G/.delta. ratio and the 
relationship between the bounding and rebounding frequency and the phase 
are shown by lines L.sub.3 in FIGS. 13(A) and 13(B). This becomes 
substantially equivalent to those obtained in measurement of actual 
lateral acceleration as illustrated by the broken lines L.sub.4. 
Therefore, according to the shown embodiment of the lateral acceleration 
projecting circuit 40, a proper projected lateral acceleration value Ge 
can be obtained with incorporating the factor of responding ability to 
bounding and rebounding frequency. 
Accordingly, by inputting the projected lateral acceleration value Ge to 
the control circuit 30 set forth above, substantially the same lateral 
acceleration dependent rolling suppressive or anti-rolling suspension 
control can be performed. In addition, since the steering angle sensor and 
the vehicle speed sensor are provided irrespective of the suspension 
control and the signals thereof can be used commonly to other vehicle 
equipment control, the cost for forming the suspension control can be 
lowered as that can neglect the lateral acceleration sensor. Furthermore, 
as set forth above, since the shown embodiment can avoid the influence of 
the lateral acceleration caused by vehicular rolling. Furthermore, since 
the lateral acceleration dependent anti-rolling suspension control can be 
performed with an OPEN LOOP control system without utilizing FEEDBACK 
control system, self-excited hunting in the control system can be 
successfully avoided. 
It will be appreciated that, when the lateral acceleration sensor is used 
for detecting lateral acceleration exerted on the vehicle, the lateral 
acceleration sensor will respond to the kinematic action of the portion of 
the vehicle body where the lateral acceleration sensor is provided. This 
action of the portion of the vehicle body is caused when the lateral 
acceleration is exerted and the vehicular rolling is indeed caused. 
Therefore, the control system is constituted as FEEDBACK control system as 
shown in FIG. 14. In such FEEDBACK control system, detecting of the 
lateral acceleration delays. In order to compensate delay in detecting the 
lateral acceleration, rise sensitivity of the FEEDBACK system becomes 
necessary. For increasing the sensitivity of the FEEDBACK control system, 
the sensitivity of the lateral acceleration sensor has to be increased. 
This will cause instability of the FEEDBACK control system due to hunting 
caused in the sensor. 
In addition, due to delay of detection of the lateral acceleration, the 
factor of lateral acceleration generated by rolling action of the vehicle 
body is incorporated in the detected lateral acceleration. 
According to the shown embodiment, since the lateral acceleration to be 
exerted on the vehicle body is projected in terms of the vehicle speed and 
the steering angular displacement, no FEEDBACK loop becomes necessary to 
successfully avoid hunting of the lateral acceleration. On the other hand, 
since the lateral acceleration can be projected before the vehicular 
rolling occurs actually, responsibility of the suspension control becomes 
substantially high. 
On the other hand, in case of the aforementioned second embodiment, the 
projected lateral acceleration value becomes greater than that of the 
actual value when lateral slip occurs on the vehicle. This can be solved 
by additionally providing the lateral acceleration sensor to the control 
system employing the lateral acceleration projecting circuit 40 of FIG. 
10. In order to detect lateral slip of the vehicle, the projected 
acceleration value Ge is compared with the measured acceleration value G 
output from the lateral acceleration sensor. An absolute value of Ge/G is 
checked to judge whether lateral slip occurs or not. When the absolute 
value is close to "1", judgement is made that lateral slip does not occur. 
In this case, the projected lateral acceleration value derived by the 
lateral acceleration projecting circuit 40 is used for suspension control. 
On the other hand, when the absolute value is substantially smaller than 
"1", judgement is made that lateral slip occurs. Then, damping rates 
.zeta..sub.1 and .zeta..sub.2 and natural frequencies .omega..sub.1 and 
.omega..sub.2 are corrected to derive the projected lateral acceleration 
with the corrected damping rates and natural frequencies. 
FIG. 16 shows the third embodiment of the active suspension control system 
according to the present invention. In this embodiment, pitching 
suppressive active suspension control is performed utilizing a pitching 
acceleration sensor 50. 
As seen from FIG. 16, the pitching acceleration sensor 50 is connected to a 
control circuit 30 which incorporates a pitching acceleration dependent 
suspension control signal generator circuit 51 and an inverter 52. The 
pitching acceleration sensor 51 is responsive to the vehicular pitching 
acceleration which causes lowering of the front and lifting up of the rear 
end, the pitching acceleration indicating signal value becomes positive. 
On the other hand, the pitching acceleration indicating signal value 
becomes negative when the pitching acceleration causes lowering of the 
rear end and lifting up at the front end. The pitching acceleration 
dependent suspension control signal generator circuit 51 is also connected 
to a gain value setting circuit 52 to receive therefrom a gain control 
signal. The pitching acceleration dependent suspension control signal 
generator circuit 51 amplifies a pitching acceleration indicating signal 
from the pitching acceleration sensor 50, with a given gain value Kx. The 
pitching acceleration dependent suspension control signal generator 
circuit 51 is connected to the solenoids 22 of the front-left and 
front-right pressure control valves 18. On the other hand, the pitching 
acceleration dependent suspension control signal generator circuit 51 is 
connected to the solenoids 22 of the rear-left and rear-right pressure 
control valves 18 via the inverter 53. Therefore, to the solenoids 22 of 
the rear-left and rear-right pressure control valves 18, a pitching 
acceleration dependent suspension control signal with opposite polarity to 
that applied to the solenoid valve 22 in the front-left and front-right 
pressure control valves 18 is applied. 
When vehicular pitching occurs, the pitching acceleration indicating signal 
is produced by the pitching acceleration sensor 50. The pitching 
acceleration dependent suspension control signal generator 51 amplifies 
the input pitching acceleration indicating signal with a given gain value 
Kx which is determined according to the gain control signal from the gain 
value setting circuit 52 to output the amplified signal as the pitching 
suppressive suspension control signal. As set forth the polarity of the 
pitching suppressive control signal to be input to the rear-left and 
rear-right pressure control valves 18 are opposite to that for the 
front-left and front-right pressure control valves 18. Therefore, when one 
of the front and rear suspension systems act to suppress bounding motion 
at the corresponding portion of the vehicle, the other suspension systems 
act for suppressing rebounding motion of the corresponding portion of the 
vehicle body. 
For example, if noise-dive occurs by braking operation, the front end 
position of the vehicle body is lowered and the rear end position is 
lifted up. In response to this, the positive value of the pitching 
acceleration indicating signal is input to the pitching acceleration 
dependent suspension control signal generator circuit 51. The pitching 
acceleration dependent suspension control signal generator circuit 51 
amplifies the received pitching acceleration indicating signal with a 
given gain Kx to output the suspension control signal. Since the positive 
value of the pitching suppressive suspension control signal is input to 
the front-left and front-right pressure control valves 18, the fluid 
pressure in the hydraulic cylinders 15A of the front-left and front-right 
suspension systems is increased to suppress lowering of the front end. On 
the other hand, since the negative value of the pitching suppressive 
suspension control signal is input to the rear-left and rear-right 
pressure control valves 18, the fluid pressure in the hydraulic cylinders 
15A of the rear-left and rear-right suspension systems is increased to 
suppress lifting up of the rear end. 
As will be appreciated, by adding the roll-damping factor, rolling speed 
dependent roll-stabilization becomes possible. This will be advantageous 
to suppress rapid change of the vehicular attitude in vehicular rolling to 
increase roll-stability of the vehicle. 
Though the embodiments have been disclosed for suppressing the attitude 
change of the vehicle body by adjusting the vehicle height, it should be 
possible to adjust only damping characteristics in response to the 
bounding and rebounding strokes of the hydraulic cylinder. Furthermore, 
though the stroke sensor comprises the potentiometer in the illustrated 
embodiment, any appropriate sensors, such as a differential transducer, 
ultrasonic distance sensor and so forth, may be applicable. In addition, 
the hydraulic cylinder may be replaced with any appropriate suspension 
element, such as pneumatic cylinder, hydropneumatic cylinder and so forth.