Actively controlled suspension system for automotive vehicles with vehicular speed dependent variable damping characteristics

Abstract: A suspension control system introduces capability of automatically adjusting sensibility of attitude change depending upon a vehicular speed with maintaining capability of manual adjustment of the sensibility for the vehicular attitude change. The actively controlled suspension system employs a pressure control valve (1 which is operable in response to a control signal and controls fluid pressure in the fluid chamber of a hydraulic cylinder (11) as a replacement of the conventional hydraulic shock absorber. The pressure control valve is associated with a controller (30) which detects bounding and rebounding acceleration and stroke speed in bounding and rebounding motion for detecting vehicular attitude change and deriving a control signal to operate the pressure control valve to suppress attitude change. The controller is response to a vehicle speed indicative parameter for adjusting sensitivity of attitude change and whereby adjusting suspension control characteristics.

Description: The present invention relates to an actively controlled suspension system for automotive vehicles as it serves for suppressing attitude changes of the vehicle body. A system according to the preamble of claim 1 is known from the patent document US-A-3,124,368. The remarkable features of this system are that a sensor for monitoring bounding/rebounding acceleration is fixed at said vehicle body close to each wheel and that the signal from each of said sensors is integrated and amplified to provide for a respective control signal for each corresponding suspension assembly. The document EP-A-0 114 757 describes an actively controlled suspension system in which the control signals, more specifically, the demanded displacements of the suspension assemblies, can be modified by signal representing vehicle speed and acceleration. In this system, two acceleration sensors are located at the nominal center of gravity of the vehicle to measure lateral and longitudinal accelerations. It is the object of the present invention to provide an actively controlled suspension system which provides for automatic adjustment of suspension control characteristics in dependence of vehicular driving conditions. The system of the present invention is defined by the features of appended claim 1. This system differs from the above firstly mentioned system in that the integrated bounding/rebounding indicative signals are multiplied with a variable gain depending on the vehicle speed. This use of the vehicle speed to modify control signals is in contrast to the teaching of the document EP-A-0 114 757, according to which demanded displacements are modified in dependence of vehicle speed and acceleration. Systems which are similar to the one of the appended claim 1 are described in the not prepublished documents EP-A-0 249 209 and EP-A-0 283 004. The system of the first just mentioned document does not comprise a bounding/rebounding sensor means at each wheel. The system of the second document does not comprise an integration means to integrate the bounding/rebounding indicative signal. In the drawings: Fig. 1 is a diagrammatical illustration showing the overall construction of the preferred embodiment of an actively controlled automotive suspension system according to the present invention; Fig. 2 is a section of one example of a hydraulic shock absorber in combination with a hydraulic pressure control valve associated with the shock absorber for adjusting stiffness of the latter; Figs. 3(A) and 3(B) are enlarged sections showing details of the hydraulic pressure control valve of Fig. 2, in which, Fig. 3(A) shows the valve position softening the shock absorber and Fig. 3(B) shows the valve position hardening the shock absorber; Fig. 4 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. 5 is a schematic block diagram of the preferred embodiment of the control section in the controller of the active suspension control system according to the invention; Fig. 6 is a schematic block diagram of a feedback system in the preferred embodiment of the suspension control system of Figs. 1 to 5; Fig. 7 is an illustration of an equivalent model utilized for discussion of operation of the preferred embodiment of the actively controlled suspension system of Figs. 1 to 5; Fig. 8 is a comparative illustration of the equivalent model of a conventional passive suspension system; Fig, 9 is a graph showing vibration transmission characteristics in relation to vibration frequency; Fig. 10 is a graph showing variation of gain for determining bounding and rebounding acceleration dependent value, in relation to equivalent mass weight of the vehicle body; Fig. 11 is a graph showing variation of gain in dependence of velocity for deriving bounding and rebounding stroke speed dependent values, in relation to spring coefficient of the suspension system; Fig. 12 is a schematic block diagram of another embodiment of the control section in the controller of the actively controlled suspension system according to the invention; Fig. 13 is a graph showing variation of gain in dependence of velocity for deriving bounding and rebounding stroke speed dependent value, in relation to spring coefficient of the suspension system in the control section of Fig. 12; and Fig. 14 is a graph showing variation of gain for determining bounding and rebounding stroke speed dependent value, in relation to equivalent mass weight of the vehicle body in the control section of Fig. 12. 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 a 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 comprises a suspension member 13, such as e.g. a suspension link, a suspension arm, 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. It should be appreciated that, in the shown embodiment, the coil spring 16 is not necessary to damp dynamic kinematic energy and should resiliently support only static load to be exerted 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 either the vehicle body 12 or the suspension member 13. On the other hand, the end of the piston rod 15b is connected to the other one of these parts. The hydraulic cylinder 15A of the suspension assembly 15 is connected to a hydraulic pressure source unit 20 via a hydraulic circuit which includes a pressure control valve 18. The pressure control valve 18 is electrically operable and connected to a control unit 30. A hydraulic circuit 19 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 associated hydraulic cylinder 15. 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 it 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 branching pressure line 22B connected to a pressure line between the pressure control valves 18 and the upper fluid chamber 15d of the hydraulic cylinder 15A. A throttle valve 22V is inserted between the junction of the branching pressure line 22B and the pressure line and the low pressure accumulator 22L. The throttle valve 22V has a fixed throttling rate. Fig. 2 shows 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. The inlet port 18b, the drain port 18c and the outlet port 18d are all 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 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 constantly communicates with the outlet port 18d and thereby communicates 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 or the drain port 18c. On the other hand, 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 actuator piston 22c contacts with an actuator rod 22a of an electrically operable actuator 22 which comprises an electromagnetic solenoid. The solenoid 22 comprises a proportioning solenoid. 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 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. 3(B) to decrease the path area at the throttle of the inner end of the inlet port 18b and to open 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. 2, 3(A) and 3(B) but can be replaced with any appropriate construction. For example, the pressure control valve constructions as illustrated in European Patent First Publication 01 93 124 can also be employed. 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 a suspension control signal V₃ from the controller 30. 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. 4. Namely, when the control value V₃ as represented by the suspension control signal is zero, the pressure P at the outlet port 18 becomes P₀ determined according to a predetermined offset pressure P₀. When the suspension control signal value increases, the fluid pressure P at the outlet port 18d increases with a predetermined proportioning gain K₁. Namely, by increasing the suspension control value V₃, the actuator rod 22a is driven downwardly in Fig. 2 toward the position of Fig. 3(A) to achieve an increase of the fluid pressure with the predetermined proportioning gain K₁. The fluid pressure P at the outlet port 18d saturates at the output pressure P₂ of the pressure unit 20. On the other hand, when the suspension control signal value V₃ decreases, the pressure P decreases to zero and shifts the actuator rod 22a toward the direction to Fig. 3(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 an 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. When bounding motion occurs at the suspension member, the piston 15c of the hydraulic cylinder 15A shifts upwardly to cause an increase of the fluid pressure in the upper chamber 15d. This causes increasing 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 with 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. When rebounding motion occurs at the suspension member, the piston 15c of the hydraulic cylinder 15A shifts downwardly to cause decreasing the fluid pressure in the upper chamber 15d. This causes decreasing 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 with decreasing 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. Here, since no flow resisting element, such as orifice, throttling valve, is disposed between the fluid reservoir 21 and the drain port 18c, no damping force against the upward motion of the piston 15c in the hydraulic cylinder 15A will be produced in response to the bounding motion of the suspension member. Since the damping force exerted on the piston 15c may allow transmission of a part of bounding energy to the vehicle body to cause rough ride feeling, the shown embodiment of the suspension system may provide satisfactorily high level riding comfort by completely absorbing the bounding and rebounding energy set forth above. In order to enable the controller 30 to perform suspension control in response to change of attitude of the vehicle body, the preferred embodiment of the active suspension control system employs bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR. The bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR are respectively oriented immediately above respectively associated front-left, front-right, rear-left and rear-right wheels 14FL, 14FR, 14RL and 14RR. The bounding and rebounding acceleration sensors 29FL, 29FR, 29RR and 29RL monitor vertical motion acceleration of the associated sections of the vehicle body in bounding and rebounding directions and thus produce bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR . The signal values of the bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR are variable depending upon direction of the motion of the associated position of the vehicle body and depending upon the magnitude of the acceleration. In the shown embodiment, the bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR generate positive bounding and rebounding acceleration indicative sensor signals, the value of which is variable depending upon the magnitude of acceleration in bounding direction at a respective associated position of the vehicle body. On the other hand, the bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR produce the bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR with negative values representing the magnitude of rebounding accelerations at respective positions. The controller 30 is further connected to a vehicle speed sensor 40 which monitors vehicle speed to produce a vehicle speed indicative signal V. In order to facilitate suspension controls for respective front-left, front-right, rear-left and rear-right suspension systems 11FL, 11FR, 11RL and 11RR, independently of each other, the controller 30 is composed of front-left, front-right, rear-left and rear-right control sections 31Fl, 31FR, 31RL and 31RR, each comprising (Fig.5) a gain controlled amplifier 35, an integrator circuit 36, a second gain controlled amplifier 37 and an adder 39. The first gain controlled amplifier 35 and the integrator circuit 36 are connected to the corresponding one of the bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR. The first gain controlled amplifier 35 is also connected to a manually operable gain setting unit 32 which may be provided near a vehicle driver so as to be manually operated for adjusting the gain Km of the first gain controlled amplifier. In practice, the first gain controlled amplifier 35 may comprise a voltage controlled variable gain amplifier, the gain of which varies depending upon a gain control voltage Zm. Variation of the gain Km in relation to the gain controlling voltage signal Zm is shown in Fig. 10. Therefore, the manually operable gain setting unit 32 may comprise a voltage setting unit for adjusting the gain controlling voltage. Therefore, the corresponding one of bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR input to the gain controlled amplifier 35 is amplified with the gain Km adjusted through the manually operable gain setting unit 32 to output an amplified signal. It will be appreciated that the amplified signal output from the first gain control signal represents as bounding and rebounding acceleration indicative factor. The integrator circuit 36 is provided in parallel to the first gain controlled amplifier 35 and connected to a second gain controlled amplifier 37 which has a variable gain Kn varying depending upon the vehicle speed indicative signal value V as shown in Fig. 11. The integrator circuit 36 in each one of the control sections 31FL, 31FR, 31RL and 31RR receives the bounding and rebounding indicative sensor signal α FL , α FR , α RL and α RR from the associated one of the bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR to output an integrated value ∫α FL dt, ∫α FR dt, ∫α RL dt or ∫α RR dt, respectively. The second gain controlled amplifier 37 is connected to the vehicle speed sensor 40 to receive therefrom a voltage signal V indicative of the measured vehicle speed, as the gain controlling voltage, to vary the amplifying gain Kn. The second gain controlled amplifier 37 amplifies the integrated value ∫α FL dt, ∫α FR dt, ∫α RL dt or ∫α RR dt respectively, with the vehicle speed dependent gain to output the amplified signal. The amplified signal output from the second gain controlled amplifier has a value representative of a bounding and rebounding stroke speed indicative factor. The bounding and rebounding acceleration representative amplified signal and the bounding and rebounding stroke speed representative amplified signal of the first and second gain controlled amplifiers 35 and 37 are fed to the adder 38. Therefore, adder 38 derives an adder signal V₄ to be fed to a comparator 39 which also receives a reference signal V₁ representative of a target vehicular height. The comparator 39 derives a difference between the adder signal V₄ and the reference signal V₁ to produce a control signal V₃ having a value corresponding to the derived difference. In the practical control, while the vehicle travels smoothly or steadily on a smooth road in straight, vehicular rolling, pitching and bounding will never occur. At this condition, the bounding and rebounding acceleration indicative sensor signal values α FL , α FR , α RL and α RR of the bounding and rebounding acceleration sensors 29FL, 29FR, 29RL and 29RR are maintained at zero. This zero values of the bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR are input to respectively corresponding front-left, front-right, rear-left and rear-right control sections 31FL, 31FR, 31RL and 31RR. Since the input values are zero, the output of the first gain controlled amplifier 35 is maintained zero. Also, the output of the integrator circuit 36 is held zero. Therefore, the suspension control signals V 3FL , V 3FR , V 3RL and V 3RR output from respective adder 39 are held zero. As set forth, while the suspension control signal value is zero, the proportioning solenoid coil 22b of the proportioning solenoid 22 is energized to provide the offset pressure P₀ to the upper fluid chamber 15d of the hydraulic cylinder 15A via the outlet port 18d. As set forth, the downward force exerted to the valve spool 19 from the bias spring 22d and the upward hydraulic pressure in the pressure control chamber 18k balances in each pressure control valve 18. At this position, when relatively high frequency and small magnitude bounding and/or rebounding vibration is input from the suspension member, the valve spool 19 vibratingly shifts in a manner set forth above to absorb vibration energy. This successfully prevents the high frequency and small magnitude bounding and rebounding vibration from being transmitted to the vehicle body. Therefore, satisfactorily high riding comfort can be achieved. On the other hand, when vehicle driving behavior causes substantial change in vehicular attitude to cause rolling, pitching or bounding, respective bounding and rebounding acceleration indicative sensor signal values α FL , α FR , α RL and α RR vary. Assuming the sensor signal values of respective bounding and rebounding acceleration indicative sensor signals α FL , α FR , α RL and α RR are α fl , α fr , α rl and α rr , the output values of respective gain deriving stages 32 of the front-left, front-right, rear-left and rear-right control sections 31FL, 31FR, 31RL and 31RR become respectively (α fl x KM), (α fr x Km), (α rl x Km) and (α rr x Km). On the other hand, based on respective bounding and rebounding acceleration indicative sensor signal values α fl , α fr , α rl and α rr , the bounding and rebounding stroke speeds Vs FL , Vs FR , Vs RL and Vs RR are derived through the integrator circuits 36 and the second gain controlled amplifiers 37 in respective control sections 31FL, 31FR, 31RL and 31RR. In practice, respective bounding and rebounding stroke speeds Vs are derived as integrated valued ∫α FL dt, ∫α FR dt, ∫α RL dt and ∫α RR dt of the bounding and rebounding acceleration indicative signal values by the integrator circuits 36. In the second gain controlled amplifiers 37, the stroke speed dependent values are derived on the basis of the bounding and rebounding stroke speed indicative values Vs FL , Vs FR , Vs RL and Vs RR and the vehicle speed dependent gain Kn. The adder 38 of the respective one of front-left, front-right, rear-left and rear-right control sections 31FL, 31FR, 31RL and 31RR derives a sum value being the adder signals V₄. Each control section 31FL, 31FR, 31RL and 31RR thus derives the control signals V 3FL , V 3FR , V 3RL and V 3RR , respectively, for the respective pressure control valve 18. For example, when right-hand steering operation is performed during vehicle travel, leftward centrifugal force may be exerted on the vehicle to cause shifting of the gravity center leftwardly. As a result, the left side of the vehicle is lowered and the right side is raised to cause vehicular rolling. Therefore, the positive bounding and rebounding acceleration indicative sensor signals α FR and α RR are produced from front-right and rear-right bounding and rebounding acceleration sensors 29FR and 29RR. On the other hand, the negative bounding and rebounding acceleration indicative sensor signals α FL and α RL are output from the front-left and rear-left bounding and rebounding acceleration sensors 29FL and 29RL. The gain derivation stage 32 of respective front-right and rear-right control sections 31FR and 31RR thus output positive stroke acceleration dependent values. The gain deriving stage 32 of the front-left and rear-left control sections 29FL and 29RR derives the negative stroke acceleration dependent values. Similarly, the arithmetic stages 33 of respective control sections derive bounding and rebounding stroke speed and thereby derive the bounding and rebounding stroke speed dependent values. This bounding and rebounding stroke speed dependent values in the front-right and rear-right control sections 31FR and 31RR are positive and that of the front-left and rear-left control sections 29FL and 29RL are negative. Therefore, the suspension control signals for the front-right and rear-right suspension systems become negative, and for the front-left and rear-left suspension systems become positive. In response to the negative suspension control signals for the front-right and rear-right suspension systems, the proportioning solenoid coils 22a of the pressure control valves 18 of the front-right and rear-right suspension systems are energized at reduced magnitude to shift the spool valves for blocking the inlet ports and establishing communication between the fluid chambers 15d and the drain ports 18c. On the other hand, in response to the positive suspension control signals, the solenoid coils 22a of the pressure control valves 18 of the front-left and rear-left suspension systems are energized at increased magnitude for shifting to increase the path area for introducing the pressure into the fluid chambers 15d of the hydraulic cylinders 15A to harden the front-left and rear-left suspension systems. As will be appreciated, with the suspension control set forth above, vehicular rolling can be successfully reduced. Here, assuming that: the mass weight of the vehicle body is M; the spring coefficient of the coil spring 16 is K; the working fluid pressure in the fluid chamber 15d is P; the effective upper surface area of the piston 15c is A; the suspension control signal value is V₄ ; the target vehicle height is V₁ ; the difference between the target height V₁ and the suspension control signal value V₄ is V₃; the gain of the pressure control valve 18 is K₁; the magnitude of displacement of the suspension member is x₁; and the magnitude of displacement of the vehicle body is x₂, the equivalent circuit to the feedback system of the controller may be illustrated as shown in Fig. 6. The inertia resistance Mx₂ of the vehicle body is determined by the resistance component of the coil spring 16 and a resistance component of the hydraulic cylinder 15A and can be illustrated by the following equation (1): Mx₂ = K (x₁ - x₂) + P.A (1) On the other hand, the difference V₃ can be illustrated by the following equation (2): V₃ = V₁ - V₄ (2) In addition, the fluid pressure P of the upper fluid chamber 15d can be illustrated by the following equation (3): P = K₁ . V₃ (3) As set forth, the suspension control signal value may be illustrated by the following equation (4): V₄ = α x Km + α x Kn (4) where α generally represents bounding and rebounding acceleration and α' is ∫αdt By replacing components in the equation (1) with the components in the equations (2) to (4), the Mx₂ may be illustrated as: The aforementioned equation (5) can be modified by Laplace transformation, as the following equation (6): MS²x₂ = - (K₁.Km.A.S².x₂ + K₁.Kn.A.S.x₂ + K.x₂) + K.x₁ + K₁.AV₁ (6) Assuming the target vehicle height V₁ is zero and disregarding the lateral accelerations, the vibration transmission characteristics X₂/X₁ can be illustrated based on the above equation (6), as following equation (7): X₂/X₁ = K/ (M + K₁.Km.A).S² + K₁.Kn.A.S + K (7) The equivalent model of the foregoing equation (7) is shown in Fig. 7. The construction of the preferred embodiment of the active suspension system according to the invention will be compared with a conventional type passive suspension system as illustrated in Fig. 8. As set forth above, in the shown embodiment, the suspension control signal value is determined on the basis of the bounding and rebounding acceleration α and bounding and rebounding stroke speed Vs which is derived on the basis of the bounding and rebounding acceleration. In this case, the vibration absorbing coefficient K₁KnA of the hydraulic cylinder 15A is disposed between the equivalent mass weight (M + K₁.Km.A) of the vehicle body and stationary section, i.e. the ground. Therefore, by adjusting the fluid pressure P in the upper fluid chamber 15d, the equivalent mass weight of the vehicle body (M + K₁.Km.A) can be directly controlled. The vibration transmission characteristics X₂/X₁ with respect to the vibration frequency is shown by the curve ℓ₁ in Fig. 9. As will be seen from Fig. 8. in the conventional passive suspension system, the spring coefficient K of the suspension coil spring and damping coefficient C of the shock absorber are disposed in parallel between the mass weight of the vehicle body to support the vehicle body with forces exerted in parallel. In this case, the vibration transmission characteristics with respect to the vibration frequency ω can be illustrated by the curve ℓ₂ in Fig. 8. In Fig. 8, the curve ℓ₃ represents the vibration transmissions characteristics with respect to the vibration frequency ω of a conventional active suspension system. As will be seen from Fig. 8, the conventional passive suspension system and the conventional active suspension system had peak of vibration transmission at around the resonating frequency, i.e. 1 Hz, of the spring. To the contrary, according to the shown embodiment, vibration transmission can be substantially reduced even at the resonance point of the spring. Here, as shown in Fig. 11, the gain Kn of the second gain controlled amplifier 37 is variable depending upon the vehicle driving speed as represented by the vehicle speed indicative signal V. Namely, the gain Kn linearly increases according to increasing of the vehicle speed indicative signal value V. This implies that when the vehicle speed is low, the gain Kn is maintained small to make the bounding and rebounding stroke indicative values Vs small. This makes the suspension control signal value greater to reduce pressure in the pressure chamber of hydraulic cylinder 15A to provide soft suspension characteristics. On the other hand, when the vehicle speed is high, the gain Kn is increased to amplify the integrated value of the integrated circuit to a greater magnitude to make the hydraulic pressure in the pressure chamber of the hydraulic cylinder increased to increase stiffness of the suspension for providing driving stability. Therefore, the shown embodiment may achieve significantly improved riding comfort. On the other hand, the preferred embodiment of the active suspension control system set forth above also may achieve satisfactorily high driving stability by suppressing change of attitude of the vehicle. In addition, since the shown embodiment introduces features for varying the suspension control signal value depending upon the vehicle driving condition so that the suspension characteristics is sufficiently soft for providing satisfactory riding comfort while the vehicle is driven at relatively low speed and that sufficiently increased stiffness of the suspension can be obtained for high driving stability while the vehicle speed is high. Though the discussion is given only for suppressing vehicular rolling in response to steering operation, the shown embodiment of the active suspension system is successful in suppressing pitching and bouncing motions of the vehicle as well. Fig. 12 shows another embodiment of the control section 31 to be employed in the suspension control system according to the invention. In this embodiment, the second gain controlled amplifier 37 is connected to another manually operable gain setting unit 41 which is manually operable for setting a desired gain by the driver. The gain setting unit produces a gain controlling voltage signal Zn which is variable for varying gain Kn of the second gain controlled amplifier for providing desired bounding and rebounding stroke speed dependent suspension control characteristics as shown in Fig. 14. The gain setting unit 41 is connected to the vehicle speed sensor 40 to receive therefrom the vehicle speed indicative signal V. The gain setting unit 41 varies the gain controlling voltage signal level depending upon the vehicle speed indicative signal value V, as shown in Fig. 13. Therefore, in comparison with the former embodiment, this embodiment will provide greater flexibility in selecting the suspension control characteristics with incorporating the vehicle speed dependent feature. While the present invention has been disclosed in terms of the preferred embodiment in order to facilitate better understanding of the invention, it should be appreciated that the invention can be embodied in various ways without departing from the principle of the invention. Therefore, the invention should be understood to include all possible embodiments and modifications to the shown embodiments which can be embodied without departing from the principle of the invention set out in the appended claims. For instance, though the shown embodiments are specific ones, it may possible to apply for other suspension control systems taking algorithms different from that set forth above. For example, the invention may be applicable for any of the embodiments disclosed in the European Patent First Publication No. 0 249 290. Furthermore, the invention is applicable not only for the hydraulic suspension system but also for pneumatic suspension system, hydro-pneumatic suspension system and so forth. As will be appreciated herefrom, the present invention fulfills all of the objects and advantages sought therefor.

Claims: An actively controlled suspension system for an auto-motive vehicle, said system comprising: a) a plurality of suspension assemblies (15), each one associated to a respective wheel and including a variable pressure chamber filled with a pressurized working fluid; b) a pressure source (20) connected to said variable pressure chambers for supplying the pressurized working fluid thereto; c) a plurality of pressure adjusting means (18), each one disposed between an associated variable pressure chamber and the pressure source for adjusting fluid pressure in the variable pressure chambers; d) a plurality of first sensor means (29FL, 29FR, 29RL, 29RR), each one being fixed at said vehicle body close to a corresponding wheel, for monitoring bounding/rebounding acceleration acting on the vehicle body to provide each a bounding/rebounding acceleration indicative signal; e) a controller (30) receiving said sensor signals and supplying a control signal to each of said pressure adjusting means to adjust the fluid pressure in the variable pressure chamber of the corresponding suspension assembly for suppressing attitude changes of the vehicle body, said controller comprising a plurality of control sections (31FL, 31FR, 31RL, 31RR) each associated with a respective wheel suspension assembly, each control section comprising: e1) an integration means (36) for integrating the bounding/rebounding acceleration indicative signal (X2FL, X2FR, X2RL, X2RR), received from the corresponding first sensor means to produce a signal representing the bounding/rebounding velocity of the vehicle body; and e2) an amplifying means (37) for amplifying the signal received from said integration means (36) to provide a respective control signal; characterized in that a second sensor means (40) is provided at said vehicle for monitoring the vehicle speed, to provide a vehicle speed indicative signal (V); and said amplifying means (37) in each of said control sections (31FL, 31FR, 31RL, 31RR) has a variable gain (Kn) increasing with an increasing value of said vehicle speed indicative signal (V). A system according to claim 1, characterized in that each of said control sections (31FL, 31FR, 31RL, 31RR) comprises an adder (38) for adding a modified or un-modified value of said bounding/rebounding acceleration indicative signal (X2FL-X2RR) with the output signal of said amplifying means to produce a control parameter signal (V4) which is fed to a comparator (39), said comparator outputting a control signal (V3) for controlling said pressure adjusting means (18) based on the difference between said control parameter (V4) and a predetermined vehicular height reference value (V1). A system as set forth in one of the claims 1 or 2, wherein said controller (30) is arranged to modify the bounding/rebounding acceleration indicative signal from each of said first sensor means by a second gain (Km), said controller adding the modified values of the bounding/rebounding acceleration indicative signal and the bounding/rebounding indicative signal together to determine a control parameter and providing control signals based on the difference between the control parameter and a preselected reference value. A system as set forth in claim 3, further comprising a manually operable gain setting unit (32) which varies the second gain (Km) at a preselected rate. A system as set forth in one of the claims 1 to 4, further comprising a second manually operable gain setting unit (41) which varies said gain (Kn) for modifying the bounding/rebounding speed indicative signal at a preselected rate.

Inventors: AKATSU YOHSUKE, FIKUSHIMA NAOTO, FUKUNAGA YUKIO, HANO SUNAO, SATOH MASAHARU, AKATSU, YOHSUKE, FIKUSHIMA, NAOTO, FUKUNAGA, YUKIO, HANO, SUNAO, SATOH, MASAHARU
Applicants: NISSAN MOTOR, NISSAN MOTOR CO., LTD.
Application year: 1988
Main classification: B60G17
IPC codes: B60G17, B60G23