Actively controlled automotive suspension system with mutually independent hydraulic systems having mutually different damping characteristics for improving response characteristics in active suspension control

Abstract: An actively controlled suspension system has two hydraulic systems for generating damping force against the vibration to be transmitted between a vehicle body and a suspension member (13). One of the hydraulic systems is designed for absorbing relatively high frequency vibration generally input from the suspension member for preventing vibration from the suspension member t=from being transmitted to the vehicle body. The other hydraulic system is active on relatively low frequency of vibration generally input from the vehicle body upon change of attitude of the vehicle body. The other hydraulic system generates a fluid pressure suppressing the attitude change. A pressure control valve (18) and a fluid pressure source (20) may be incorporated in the second hydraulic system so that they may not be active when aforementioned one of the hydraulic system is active.

Description: BACKGROUND OF THE INVENTION Field of the Invention The present invention relates to a suspension system for a vehicle according to the preamble of claim 1. A suspension system of this kind is described by the EP-A-2 0 249 209 going back to a prior application. Another actively controlled suspension system is disclosed in IEEE Transactions of Industrial Electronic. This suspension system comprises a first and a second pressure adjusting means in connection with a pressure chamber of a cylinder for absorbing high frequency vibrations input from a suspension member and low frequency vibrations input from the vehicle body. A system of this kind is also disclosed by the DE-A-27 36 026. The DE-B-27 15 895 describes an aircraft suspension having shock damping means and load adjusting means in connection with a common pressure chamber of a cylinder. 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 automotive suspension system with the suspension member. The suspension coil spring generally resists against 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-stablizer 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 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 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 has been proposed in European Patent First Publication 01 93 124, published on September 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 a 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 presure 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 chamber 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 September 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 the construction, the pressurized fluid to absorb road shock and suppress attitude change of the vehicle body has been supplied from the pressure source via the pressure control valve. This loweres responsabiulity of pressure control in the hydraulic actuator. In addition since an engine-driven pump is employed for generating pressurized fluid flow, the extra engine load due to the load of the pump becomes substantial to lower the fuel economy of the engine. SUMMARY OF THE INVENTION Therefore, it is an object of the present invention to solve the problems in the prior art and to provide an active controlled suspension system which has substantially high respobability. Another object of the invention is to provide an actively controlled suspension system which will reduce reduced load for an automotive engine for generating fluid pressure for pressure control. In order to accomplish the aforementioned and other objects, an actively controlled suspension system, according to the invention employs in combination the features disclosed in claim 1. A pressure control valve and a fluid pressure source are incorporated in the second hydraulic system so they they may not be active when aforementioned one of the hydraulic system is active. Further particular embodiments are disclosed in the dependent claims. The first pressure adjusting means includes a pressure accumulator connected to the variable pressure chamber and a first flow restricting means interposed between the variable pressure chamber and the pressure accumulator for liming flow of the fluid for absorbing the vibration energy. On the other hand, the second pressure adjusting means includes a pressurized fluid source, a circuit connecting the pressrized fluid source, a valve means for adjusting flow of pressuirzed fluid, and a controller detective of attitude change of the vehicle body for operating the pressure control value for adjusting fluid pressure in the variable pressure chamber for suppressing attitude change of the vehicle body. The second adjusting means incorporates a second flow restricting means for restricting fluid flow between the variable pressure chamber and the pressure control valve, daid second flow restricting means having variable flow restriction rate variable depending upon vibration frequency so that the flow restriction rate becomes greater than that of the first flow resticting means when the vibration frequency is higher than a given value. The given value is set at a boarder frequency between a resonance frequency of the vehicle body and a resonance frequency of the suspension member as coupled with the road wheel. The controller is connected to a detector detecting change of attitude of the vehicle body to produce an attitude change indicative signal indicative of magnitude of change of attitude, and the controller derives a control value for controlling the pressure control valve and whereby controlling fluid pressure in the variable pressure chamber to produce a control signal for controlling the pressrue control valve. In the preferred embodiment, the detector comprises an acceleration sensor monitoring bounding and rebounding acceleration of the vehicle body for producing a bounding and rebounding acceleration indicative signal. The controller integrates the bounding and rebounding acceleration indicative signal value for obtaining bounding and rebounding motion indicative value to derive the control signal value based thereon. BRIEF DESCRIPTION OF THE DRAWINGS In the drawings: Fig. 1 is a diagrammatical illustration showing the overall construction of the preferred embodiment of an automotive suspension system with a control system actively controlling the suspension system according to the present invention; Fig. 2 is a diagramatic illustration of the preferred embodiment of one of the active suspension systems employed in the preferred embodiment of the actively controlled suspension system according to the invention; Fig. 3 is a section of one example of a hydraulic cylinder in combination with a hydraulic pressure control valve associated with the shock absorber for adjusting stiffness of the latter; Figs. 4(A) and 4(B) are enlarged sections showing the detail of the hydraulic pressure control valve of Fig. 3, in which, Fig. 4(A) shows the valve position softening the shock absorber and Fig. 4(B) shows the valve position hardening the shock absorber; Fig. 5 is a graph showing variation of a hydraulic pressure to be supplied to a pressure control value according to variation of drive signal voltage; Fig. 6 is a graph showing variation of worling fluid flow resistance in relation to a frequency of input vibration; and Fig. 7 is a section similar to Fig. 3 but showing a modified construction of the hydraulic pressure control valve to be employed in the actively controlled suspension system of Fig. 1. DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now 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 so-called active suspension system which adjusts worling fluid pressure in the suspension system for adjusting stiffness and damping characteristics of the suspension system in positive manner for absorbing discomfortable relatively high frequency and relatively small magnitude vibration input from the road wheels 14FL, 14FR, 14 RL 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 uspension 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 shown embodiment, the suspension coil spring 16 is not necessary to damp the bounding and rebounding kinematic energy and is required onlt to resiliently support the vehicle body on the road wheel resistanting the static load due to the mass weight of the vehicle body. In the further positive side, the suspension coil spring 16 should be weak enough not to produce 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 houisng 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 electrically operable 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 a low pressure accumulator 22L are connected between the pressure control valves 18 and the associated hydraulic cylinder 15A. The pressure source unit 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 Figs. 1 and 2, the low pressure accumulator 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 accumulator 22L. The throttle valve 22V has a fixed throttling rate to provide a predetermined flow resistance against the working fluid therethrough. The controller 30 is connected to a bounding and rebounding acceleration sensors 29FL 29FR, 29RL and 29RR. The bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR are respectively located just above the front-left, front-right, rear-left and rear-right wheels 14FL, 14FR, 14RL and 14RR to output bounding and rebounding acceleration indicative signals α FL , α FR , α RL and α RR . The controller 30 derives the suspension control signals for controlling respective pressure control valves 18 based on the bounding and rebounding acceleration indicative signals α FL , α FR , α RL and α RR . Fig. 3 shows the detailed construction of the hydraulic cylinder 15A and the pressure control valve 18. As will be seen from Fig. 3, 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 communciation line 27. The pressure control valve 18 has a valve housing 18A having the forementioned outlet port 18d, and 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. 3, the third land 19c has smaller diameter that 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 18 which is connected to the drain port 18c via a drain path 18c. An annular pressure chamber 18j is defined between the first and second lands 19a and 19b. The pressure chamber 18j is constantly communicated with the outlet port 18d and whereby communicated 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 presure 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 22c contacts with an actuator rod 22a of an electrically operable actuator 22 which comprises an electromagnetic solenoid. The solenoid 22 is a proportioning solenoid which varies magnitude of actuation of the actuator rod 22 a to determin 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. 4(A) to increase path area at a throttle constituted 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. 4(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 specified to the construction as illustrated in Figs. 3, 4(A) and 4(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 12 is herein incorporated by reference for the sake of disclosure. As seen from Fig. 3, 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 3 from the controller 30. 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 shown 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. 5. Namely, when the suspension control signal V 3 is zero, the pressure P at the outlet port 18 becomes P 0 determined according to a predetermined offset pressure P 0 . When the suspension control signal value in positive value increases, the fluid pressure P at the outlet port 18d increases with a predetermined proportioning gain K 1 . Namely, by increasing of the suspension control valve V 3 , the actuator rod 22a is driven downwardly in Fig. 3 at a magnitude toward to position of Fig. 4(A) to achieve increasing of the fluid pressure with the predetermined proportioning gain K 1 . The fluid pressure P at the outlet port 18d saturate at the output pressure P 2 of the pressure unit 20. On the other hand, when the suspension control signal value V 3 decreases, the pressure P decreases to zero to be shifting of the actuator rod 22a toward the direction to Fig. 4(B). The actuator rod 22a of the proportioning solenoid 22 is associated 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 constitutes the second hydraulic system. On the other hand, the low pressure accumulator 22L, the blanch line 22B and the throttle valve 22V constitute the first 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 C 2 against the working fluid flow therethrough. The flow resistance C 2 varies accroding to input vibration frequency which corresponds to the stroke speed of the piston 15c of the hydraulic cylinder 15A in non-linear fashion. More practically, as shown in Fig. 6, the variation characteristics of the flow resistance C 2 in the communication path 27 is parabolic characteristics in relation to the vibration frequency. Furthermore, the flow resistance value C 2 of the communication path is set smaller than the flow resistance C 1 in the first hydraulic system set forth above. when the input vibration frequency is lower than a boader frequency f 0 , 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 C 2 of the communication path is set greater than or equal to the flow resistance C 1 in the first hydraulic system set forth above, when the input vibration frequency is higher than or equal to the boader frequency f 0 . 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 boader frequency f 0 . 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 boader frequency. Since the flow resistance of the communication path 27 becomes greater than that of the first hydraulic system when the vibration is input from the suspension member. Therefore, in this case, the first hydraulic system becomes active to absorb vibration energy. On the other hand, when the vibration frequency is lower than the boader frequency f 0 during vehicular attitude change, such as vehicular rolling, pitching and bouncing and so forth, the second 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 boarder 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 C 1 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 damoping force for absorbing vibration energy so that the vibration energy is not transmitted to the vehicle body. On the other hand, when nose-dive occurs due to vehicular braking operation, the front end of the vehicle body is lowered. At this time, the vibration frequency of the pitching motion, i.e. nose-dive, of the vehivle body is normally lower than the boarder frequency f 0 . Therefore, the flow resistance of the second hydraulic system becomes lower than that of the first hydraulic system. On the other hand, lowering of the front end of the vehicle body is detected by the bounding and rebounding acceleration sensors 29FL and 29FR. The bounding and rebounding acceleration sensors 29FL and 29FR thus outputs the positive values of bounding and rebounding acceleration indicative sensor signals. The controller 30 integrates the bounding and rebounding acceleration indicative sensor signal values α FL and α FR to derive the bounding and rebounding stroke speed indicative value V SFL and V SFR . Based on the bounding and rebounding stroke speed indicative value V SFL and V SFR , the controller outputs the suspension control signal having values corresponding thereto. By the suspension control signal, the solenoide coils 22b of the proportioning solenoids 22 are energized at greater magnitude to cause downward movement of the actuation rods 22a. This increases induction rate of the pressurized fluid from the pressurized fluid source unit 20 to the fluid chamber 15d of the hydraulic cylinder 15A. This increase stiffness of the suspension system to suppress lowering motion of the front end of the vehicle body. Therefore, nose-dive in response to the vehicular braking operation can be successfully prevented. On the other hand, when winding up occurs due to abrupt acceleration of the vehicle, the front end of the vehicle body is lifted up. In response to rising of the front end of the vehicle body, the bounding and rebounding acceleration sensors 29FL and 29FR outputs the nagative values of bounding and rebounding acceleration indicative signals α FL and α FR . Therefore, the bounding and rebounding stroke speed indicative valuies V SFL and V SFR based on the bounding and revbounding acceleration indicative sensor signals. This bounding and rebounding stroke speed indicative value is output to the solenoid coils 22b as the suspension control signals. Therefore, the energization magnitude of the solenoid coils are reduced to cause upward shifting of the actuation rods 22a. This reduces supply rate of the working fluid and causes drainig of the working fluid in the fluid chamber 15d. As a result, the front-left and front-right suspension systems are softened to suppress upward motion of the front end of the vehicle body. On the other hand, when the vehicle body causes bouncing by running over a projection or dent on the road, the balance between the spring force of the bias spring 22d and the fluid pressure in the presusre control chamber 18k is distroyed due to variation of the fluid pressure at the outlet port 18d. This causes shifting of the valve spool 19 to absorb the vibration energy for suppressing bouncing action of the vehicle body. As will be appreciated herefrom, the shown embodiment avoid consumption of the fluid pressure to be generated in the pressuirzed fluid source unit 20 for absorbing relatively high frequency vibration such as road shock. This will reduce necessity of driving of the engine-driven fluid pump in the pressurized fluid source unit to reduce the engine load. This will improves fuel economy in driving the vehicle. Though the shown embodiment provides the flow resistance difference between the first hydraulic system and the second hydraulic system by adjusting the length and path area of the communication path 27. However, difference of the working fluid flow characteristics as required can be obtained in various ways. Therefore, the invention should not be limited to provide the flow resistance difference between the first and second hydraulic systems in the construction set forth above. One of the examples of the modified hydraulic circuit for providing the equivalent flow resistance dfference has been illustrated in Fig. 7. In the modification of Fig. 7, an orifice 18h is formed in the pilot path 18g connecting the pressure control chamber 18k and the outlet port 18c. The throttling rate in the orifice is set to provide fluid flow characteristics to increase flow resistance in the first hydraulic system according to increasing of the vibration frequency. With the construction, by providing appropriate throttling rate in the orifice 18h, the fluid flow characteristics to increase the flow resistance C 2 across that C 1 of the second hydraulic system when the vibration frequency increases across the boarder frequency f 0 . As will be appreciated herefrom, according to the shown embodiment, since the first hydraulic system is active for absorbing the high frequency vibration, such as road shock, responsability against input of the shock is substantially high to successfully prevent the vibration to be transmitted to the vehicle body. On the other hand, since the second hydraulic system is active only upon the low frequency vibration which is caused in response to rolling, pitching, bouncing and so forth, wasting of the engine driving torque for driving the fluid pump in the pressurized fluid source unit can be minimized. It should be noted that, though the present invention has been disclosed in terms of the preferred embodiment of the invention which has been directed to an active suspension system employing a hydraulic cylinder, the invention should be applicable for any type of suspension systems which is adjustable of damping characteristics, such as pneumatic suspension system, hydropneumatic suspension system and so forth.

Claims: An actively controlled suspension system for an automotive vehicle comprising: a suspension assembly (11) interposed between a vehicle body (12) and a suspension member (13) rotatingly supporting a road wheel (14), said suspension assembly including a variable pressure chamber (15d) connected to a pressure source (20), the pressure in the variable pressure chamber being variable in response to vibration input of said suspension assembly, a first pressure adjusting means (22L,22V,22B) associated with the variable pressure chamber (15d) and responsive to vibration input from said suspension member (13) for adjusting pressure in said variable pressure chamber (15d) for absorbing vibration energy input in said suspension assembly (11), and a second pressure adjusting means (22,18,19,30) associated with the variable pressure chamber (15d) and responsive to vibration input from said vehicle body (12) for adjusting pressure in said variable pressure chamber (15d) for suppressing change of vehicular attitude, characterized in that both said first and second pressure adjusting means are formed by a unique pressure control valve assembly disposed within a fluid circuit connecting said variable pressure chamber (15d) and said pressure source (20), said first pressure adjusting means includes a pressure accumulator (22L) connected to said variable pressure chamber (15d) and a first flow restricting means (22V,22B) interposed between said variable pressure chamber and said pressure accumulator for limiting flow of said fluid for absorbing said vibration energy, said second pressure adjusting means (22,18,19,30) includes a pressurized fluid source (20,21), a valve means (18) for adjusting flow of pressurized fluid, a controller (30) detective of attitude changes of the vehicle body for operating said valve means (18) for adjusting fluid pressure in said variable pressure chamber (15d) for suppressing attitude changes of the vehicle body, a second flow restricting means (27, 18h) for restricting flow between said variable pressure chamber (15d) and said valve means (18), said second flow restricting means having a variable flow restriction rate variable depending upon vibration frequency so that said flow restriction rate becomes greater than that of said first flow restricting means when said vibration frequency is higher than a given value (fo), said given value (fo) being set at a border frequency between a resonance frequency of the vehicle body (12) and a resonance frequency of said suspension member (13) as coupled with said road wheel (14). An actively controlled suspension system as claimed in claim 1, characterized by a detector detecting a change of attitude of the vehicle body, said detector comprising an acceleration sensor (29) monitoring bounding and rebounding acceleration of the vehicle body for producing a bounding and rebounding acceleration indicative signal. An actively controlled suspension system as claimed in claim 2, characterized by a controller (30) integrating said bounding and rebounding acceleration indicative signal value for obtaining a bounding and rebounding motion indicative value to derive said control signal value based thereon.

Inventors: AKATSU YOHSUKE, FUKUNAGA YUKIO, FUKUSHIMA NAOTO, HANO SUNAO, AKATSU, YOHSUKE, FUKUNAGA, YUKIO, FUKUSHIMA, NAOTO, HANO, SUNAO
Applicants: NISSAN MOTOR, NISSAN MOTOR CO., LTD.
Application year: 1987
Main classification: B60G17
IPC codes: B60G17, F16F9, B60G23