ANTI-ROLLING CONTROL FOR ACTIVELY CONTROLLED AUTOMOTIVE SUSPENSION SYSTEM WITH ENHANCED ROLLING DETECTING ABILITY

Abstract: An actively controlled suspension system employs a lateral acceleration sensor for monitoring lateral acceleration exerted on the vehicular body. The lateral acceleration is shifted fronwardly from a gravity center of the vehicle body at the neutral position or non-rolling condition. The magnitude of fronward offset of the lateral acceleration sensor is so selected at to cancel influence of phase delay when the lateral acceleration sensor is oriented at the position of the neutral gravity sensor.

Description: The present invention relates to an actively controlled suspension system for an automotive vehicle according to the preamble of claim 1. The EP-A-2 049 246 discloses an active vehicle suspension system corresponding to the preamble of claim 1. European Patent First Publication No. 0 283 004 discloses an actively controlled suspension system including an anti-roll control loop which is responsive to lateral acceleration to be exerted on a vehicle to adjust suspension characteristics. The control loop employing suspension system is variable of response characteristics to vehicular rolling depending upon the vehicle speed for varying anti-rolling suspension control characteristics. In a range of the vehicle speed in which the greater loop gain of the control loop causes significant and unacceptable level of self-induced lateral vibration, the gain is varied to be smaller than that in other vehicle speed range. Namely, since the greater gain may cause self-induced lateral vibration in the low speed range, the gain main be adjusted to a smaller value while vehicle speed is lower than a given value. Also, European Patent First Publication No. 0 285 153 discloses an actively controlled automotive suspension system employing a factor representative of self-induced attitude change creative factor as correction factor in attitude regulating suspension control. Suspension characteristics is controlled on the basis of rolling and/or pitching energy representing parameter. The correction factor representative of self-induced attitude change is derived on the basis of the vehicular attitude change velocity for canceling the self-induced rolling and/or pitching. By canceling the self-induced rolling and/or pitching creative component from the rolling and/or pitching acceleration factor, higher rolling and/or pitching response characteristics can be provided so as to satisfactorily high driving stability and riding comfort. On the other hand, Japanese Patent First Publication 62-295714, which is assigned to the common assignee to the present invention, discloses one typical example of prior proposed actively controlled suspension systems for performing anti-rolling and/or anti-pitching suspension controls. The disclosed system employs a lateral acceleration sensor and a longitudinal acceleration sensor for monitoring vehicular attitude change magnitude indicative parameters for performing attitude change suppressive control. As will be appreciated, the lateral acceleration sensor employed in the actively controlled suspension system monitors the lateral force exerted on the vehicular body and detects the possible magnitude of vehicular rolling. The system controls suspension characteristics of respective suspension systems suspending vehicular body on respective front-left, front-right, rear-left and rear-right wheels. In the foregoing publication, a singular lateral acceleration sensor for providing common anti-rolling control parameter for controlling suspension characteristics of respective suspension systems. When single lateral acceleration sensor is to be mounted on the vehicle body, it is preferred to locate the sensor at or nearby the gravity center. When the lateral acceleration sensor is provided at the gravity center, lag factors, such as mechanical lag in a control valve, phase delay in a control system, will substantially affect for response characteristics of anti-rolling control. In order to improve this, it has been proposed to locate the lateral acceleration sensor at an orientation frontwardly shifted from the gravity center. In such case, a difficulty is encountered in that, since the front portion of the vehicle body is generally occupied by an automotive engine and a power transmission, an appropriate position to install the sensor is difficult to find out. Furthermore, for avoiding influence of rolling motion of the vehicular body in monitoring the lateral acceleration, the height level to install the lateral acceleration sensor has to be as close as possible to vehicular body rolling axis. This further makes it difficult to find out an appropriate position to install the sensor. In addition, as will be appreciated, the gravity center of the vehicular body shifts from an initial position depending upon load condition, such as increasing or decreasing of luggages or passengers, anti-rolling control characteristics tend to be varied in relation to the position of the gravity center. This makes the vehicle driving characteristics unstable due to variation of anti-rolling characteristics. In view of the difficulties in the prior proposed systems, it is an object of the present invention to provide an actively controlled suspension system which can perform precise anti-rolling suspension control with compensation of phase delay caused in a control system. In order to accomplish the object of the invention, a control system according to the invention is characterized by the characterizing features of claim 1. Further preferably, the hydraulic cylinder and the pressure control valve means is provided for each of suspension systems respectively associated with the front and rear wheels, and the controller means derives the suspension control signal for the pressure control valve means associated with the suspension systems of the front wheels and the rear wheels on the basis of the sensor signal. The frontward offset magnitude of the acceleration sensor is derived so as to compensate phase delay of control system. The frontward offset magnitude of the acceleration sensor relative to the gravity center is derived so that the acceleration sensor is additionally subject yawing force in a magnitude compensating the phase delay. The present invention will be understood more fully from the detailed description given herebelow and from the accompanying drawings of the preferred embodiment of the invention, which, however, should not be taken to limit the invention to the specific embodiment but are for explanation and understanding only. In the drawings: Fig. 1 is an illustration of the overall construction of the preferred embodiment of an actively controlled suspension system according to the present invention; Fig. 2 is a sectional view of a pressure control valve employed in the preferred embodiment of the actively controlled suspension system of Fig. 1; Fig. 3 is another embodiment of the pressure control valve to be employed in the preferred embodiment of the actively controlled suspension system of Fig. 1 Fig. 4 is a circuit diagram of a hydraulic circuit to be employed in the preferred embodiment of the actively controlled suspension system according to the present invention; Fig. 5 is a chart showing relationship between an electric current value of a control signal to be output by a control unit and a hydraulic pressure in a pressure chamber of a hydraulic cylinder; Fig. 6 is a schematic and illustrative block diagram of a control circuit to be employed in the preferred embodiment of the actively controlled suspension system according to the invention; Fig. 7 is a chart showing relationship between lateral acceleration caused by steering operation and output signal values of the lateral acceleration sensors; Fig. 8 is a block diagram of the preferred construction of a suspension control system which is employed in the preferred embodiment of the actively controlled suspension system of Fig. 1; Fig. 9 is a circuit diagram of a driver circuit which is applicable for the suspension control system of Fig. 8; Fig. 10 is an illustration showing preferred position of the vehicular body for monitoring the lateral acceleration; Figs. 11(a) and 11(b) are graphs showing variation of gain and phase in response transferring coefficient of output of lateral acceleration sensor with taking vehicular speed as parameter; Figs. 12(a) and 12(b) are graphs showing variation of gain and phase in response transferring coefficient of output of lateral acceleration sensor with taking a distance between the installing position and the gravity sensor as parameter; and Figs. 13(a) and 13(b) are graphs showing response transferring coefficient of rolling rate versus actual steering angle. Referring now to the drawings, particularly to Fig. 1, the preferred embodiment of an actively controlled suspension system, according to the present invention, is designed to generally perform suspension control for suppressing relative displacement between a vehicle body 10 and suspension members 24FL, 24FR, 24RL and 24RR provided in front-left, front-right, rear-left and rear-right suspension mechanisms 14FL, 14FR, 14RL and 14RR and rotatably supporting front-left, front-right, rear-left and rear-right wheels 11FL, 11FR, 11RL and 11RR, which suspension members will be represented by the reference numeral ''24'' as generally referred to, and suspension mechanisms will be represented by the reference numeral ''14'' as generally referred to, and whereby suppressing vehicle body attitude change. Respective front-left, front-right, rear-left and rear-right suspension mechanisms 14FL, 14FR, 14RL and 14RR include hydraulic cylinders 26FL, 26FR, 26RL and 26RR, which hydraulic cylinder will be hereafter represented by the reference numeral ''26'' as generally referred to. Each of the hydraulic cylinder 26 is disposed between the vehicle body 10 and the suspension member 24 to produce a damping force for suppressing relative displacement between the vehicle body and the suspension member. The hydraulic cylinder 26 has a cylinder body 26a. The cylinder body 26a defines a working fluid chamber therein. A piston 26c is disposed within the working fluid chamber to divide the working fluid chamber into a upper working chamber 26d and a lower reference pressure chamber 26e. The working chamber 26d and the reference pressure chamber 26e are communicated each other via an orifice defined through the piston 26c. The piston 26c is connected to the associated one of the suspension members 24FL, 24FR, 24RL and 24RR via a piston rod 26b. A suspension coil spring 36 is provided in parallel to each of the hydraulic cylinder 26. The suspension coil spring 36 employed in the shown type of the suspension is not required a resilient force sufficient for damping relative displacement between the vehicle body and the associated suspension member. The working chamber 26d of each hydraulic cylinder 26 is connected to a hydraulic pressure source unit 16 via a pressure supply line 52, a pressure drain line 54, a pressure control valve units 28FL, 28FR, 28RL and 28RR and a pressure control line 38, which pressure control valve units will be hereafter represented by the reference numeral ''28'' as generally referred to, and a pressure line 38. As seen from Figs. 1, the pressure line 38 is connected to a fluid path 26g defined through the piston rod 26b and the piston 26c. The working chamber 26d of the hydraulic cylinder 26 is further communicated with a pressure accumulator 34 via an orifice 32. Other pressure accumulators 18 are provided in the pressure supply line 52 for accumulating the excessive pressure generated by the pressure source unit 16. The pressure control valve unit 28 comprises a proportioning valve and is designed to be controlled by an electric control signal for varying valve position according to variation of current value of the control signal. Generally, the pressure control valve unit 28 controls magnitude of introduction and draining of the pressurized working fluid into and from the working chamber 26d for adjusting the pressure in the working chamber for setting the damping mode of the hydraulic cylinder 26. To control the valve position of the pressure control valve unit 28, a control unit 22 which comprises a microprocessor, is provided. The control unit 22 is connected to various sensors which monitor vehicle body attitude change representative parameters to produce sensor signals. The sensors may include vehicular height sensors 21 for monitoring stroke of relative motion between the vehicle body and the suspension members 24 to produce vehicle height indicative sensor signals, and lateral acceleration sensor 23 which are designed for monitoring lateral acceleration exerted on the vehicle body to produce lateral acceleration indicative signals. As will be understood from the discussion given later, the lateral acceleration sensor 23 is longitudinally oriented at a position shifted frontwardly. While the shown embodiment employs the vehicle height sensor as the vehicular attitude change indicative parameter, the similar vehicular height indicative parameter can be monitored by a vertical acceleration sensor which monitors vertical acceleration exerted to the vehicle body, to produce vertical acceleration indicative signal for bouncing control for suppressing bouncing and rebouncing motion of the vehicle body. In addition, a longitudinal acceleration sensor for monitoring longitudinal acceleration exerted on the vehicle body as vehicular pitching representative parameter may be employed for anti-pitching and anti-rolling control. Furthermore, other sensors, such as a vehicular speed sensor, steering angle sensor and so forth which monitor vehicular driving condition affecting vehicular attitude may also be employed for performing various suspension control. Fig. 2 shows the detailed construction of the pressure control valve unit 28 to be employed in the shown embodiment of the actively controlled suspension system set forth above. The pressure control valve 28 comprises a valve housing 42 and a proportioning solenoid 43. The proportioning solenoid 43 is electrically connected to the control unit 22. The valve housing 42 defines a valve bore which is separated into a valve chamber 42L and a control chamber 42U by means of a partitioning member 42A. The partitioning member 42A is formed with a communication hole 42a. Above the communication hole 42a is defined the control chamber 42U. The valve chamber 42L and the control chamber 42U are aligned to each other across the communication hole 42a. Beneath the communication hole 42a and adjacent the top of the valve chamber 42L, a stationary throttling orifice defining member 44 is provided. The throttling orifice defining member 44 is formed with a fixed throttling rate of orifice. The throttling orifice defining member 44 defines with the partitioning member 42A a pilot chamber P. A valve spool 48 is thrustingly or slidingly disposed within the valve chamber 42L. The valve spool 48 defines an upper feedback chamber FU between the top end thereof and the throttling orifice defining member 44. The valve spool 48 also defines a lower feedback chamber FL between the lower end thereof and the bottom of the valve chamber 42L. Offset springs 50A and 50B are disposed within the upper and lower feedback chambers FU and FL, which offset springs exert spring force to the valve spool 48 for resiliently restricting movement of the latter. The valve chamber 42L is communicated with an inlet port 42i, a drain port 42o and the communication port 42n which are defined through the valve housing 42. The inlet port 42i is connected to the pressure unit 16 via a supply line 52. On the other hand, the drain port 42o is connected to the pressure unit 16 via the drain line 54. The valve spool 48 is formed with an upper land 48b and a lower land 48a. The upper and lower lands 48b and 48a define therebetween an annular pressure chamber 48c. The valve spool 48 is formed with a pilot path communicating the pressure chamber 48c with the lower feedback chamber FL. A poppet valve member 56 is disposed within the control chamber 42U for thrusting or sliding movement therein. The poppet valve member 56 has a valve head opposing to the communication hole 42a. The poppet valve member 56 is operably associated with the proportioning solenoid 43 which comprises a plunger 58 which has a plunger rod 58A. The lower end of the plunger rod 58A of the plunger 58 opposes to the top end of the poppet valve 56. The poppet valve member 56 is driven by the plunger 58 to control the path area in the communication hole 42a according to the position of the plunger rod 58A. Therefore, the poppet valve member 56 adjusts the path area of the communication hole 42a to control fluid pressure to be introduced in the pilot chamber P. Poppet valve member 56 separates the control chamber 42U into upper and lower control chambers. In order to control the position of the poppet valve 56 for adjusting the pilot pressure in the pilot chamber P, a solenoid coil 60 is provided for energizing the deenergizing to cause axial shift of the plunger rod 58A. By adjusting the fluid pressure in the pilot chamber P, the pressure in the upper feedback chamber FU is adjusted to exert an axially driving force to the valve spool 48 to cause axial shift. By this, selective fluid communication between the inlet port 42i, the drain port 42o and the communication port 42n can be established to adjust the fluid pressure at the communication port 42n. Since the pressure at the communication port 42n is equal to the fluid pressure in the working chamber 26d of the pressure cylinder 26, the damping force created by the pressure cylinder can be adjusted. The inlet port 42i is also connected to the pilot chamber P via a fluid path 42s. On the other hand, the drain port 42o is connected to control chamber 42U via a fluid path 42t. In order to control the pressure at the communication port 42n, a control current I is applied to the solenoid coil 60. Fig. 3 shows alternative embodiment 28' of the pressure control valve unit 28 to be employed in the preferred embodiment of the actively controlled suspension system, according to the invention as well as the alternative embodiment 26' of the hydraulic cylinder 26. As will be seen from Fig. 3, the hollow cylinder housing 115a of cylinder 26' is formed with a port 115f communicating the upper fluid chamber 15d to an outlet port 118d of the pressure control valve 28' via a communication line 127. Though Fig. 3 does not show clear construction, the lower fluid chamber 115e is defined as an enclosed space and is filled with the viscous working fluid. The pressure of the working fluid in the lower fluid chamber 115e at an initial position of the piston 115c serves as a reference pressure and per se serves as resistance for downward movement of the piston. The pressure control valve 28' has a valve housing 118 having the aforementioned outlet port 118d, an inlet port 118b and a drain port 118c. Respective inlet port 118b, the drain port 118c and the outlet port 118d are connected to a valve bore 118a defined within the valve housing 118A. A valve spool 119 is disposed within the valve bore 118a for thrusting movement therein. The valve spool 119 has first, second and third lands 119a, 119b and 119c. As will be seen from Fig. 3, the third land 119c has smaller diameter than that of the first and second lands 119a and 119b. The third land 119c defines a fifth pressure control chamber 118h which is connected to the drain port 118c via a drain path 118f. An actuator piston 122c is also disposed within the valve bore 118a. The actuator piston 122c opposes the second land 119b in spaced apart relationship to define a second pressure control chamber 118i which is connected to the drain port 118c via a drain path 118e. An annular pressure chamber 118j is defined between the first and second lands 119a and 119b. The pressure chamber 118j is constantly communicated with the outlet port 118d and whereby communicated with the upper fluid chamber 115d. On the other hand, the pressure chamber 118j shifts according to shifting of the valve spool 119 to selectively communicate with the inlet port 118b and the drain port 118c. On the other hand, an pressure control chamber 118k is defined between the first and third lands 119a and 119c. The pressure control chamber 118k is in communication with the outlet port 118d via a pilot path 118g. A bias spring 122d is interposed between the actuator piston 122c and the valve spool 119. The actuator piston 122c contacts with an actuator rod 122a of an electrically operable actuator 122 which comprises an electromagnetic solenoid. The solenoid 122 comprises a proportioning solenoid. In order to increase the supply pressure of the working fluid, the spool valve 119 is shifted to the position to increase path area at a throttle constituted at the inner end of the inlet port 118b by means of the land 119a of the spool valve 119. On the other hand, in order to decrease the supply pressure of the working fluid, the spool valve is shifted to the position to decrease the path area at the throttle of the inner end of the inlet port 118b and opens the drain port 118 which is normally blocked by means of the land 119b of the spool valve. As seen from Fig. 3, the proportioning solenoid 122 comprises the actuator rod 122a and a solenoid coil 122b. The solenoid coil 122b is energized by suspension control signal from the control unit. In the shown embodiment of the pressure control valve, the working fluid pressure P at the outlet port 118d is variable according to the predetermined variation characteristics. Namely, when the control value represented by the suspension control signal is zero, the pressure at the outlet port 118 becomes an initial pressure determined according to a predetermined offset pressure. When the suspension control signal value is a positive value and increases, the fluid pressure at the outlet port 118d increases with a predetermined proportioning rate. Namely, by increasing of the suspension control value, the actuator rod 122a is driven downwardly in Fig. 3 toward a position to achieve increasing of the fluid pressure with the predetermined proportioning rate. The fluid pressure at the outlet port 118d saturates at the output pressure of the pressure unit. On the other hand, when the suspension control signal value decreases, the pressure decreases to zero by shifting of the actuator rod 122a. The actuator rod 122a of the proportioning solenoid 122 is associated with the actuator piston 122c. Contact between the actuator rod 122a and the actuator piston 122c can be maintained by the resilient force of the bias spring 122d which normally biases the actuator piston toward the actuator rod. On the other hand, the spring force of the bias spring 122d is also exerted on the valve spool 119 to constantly bias the valve spool downwardly in Fig. 3. The valve spool 119 also receives upward hydraulic force from the pressure control chamber 118k. Therefore, the valve spool 119 is oriented at the position in the valve bore at the position where the downward bias of the bias spring 122d balances with the upward hydraulic force of the pressure control chamber 118k. Fig. 4 is a circuit diagram of a hydraulic circuit including the pressure source unit 16 and the working fluid chamber 26d of the hydraulic cylinder 26. Fig. 4 shows detailed circuit construction of the preferred embodiment of the hydraulic system to be employed in the suspension control system according to the invention. As set forth, the pressure source unit includes the pressure unit 16 which comprises a fluid pump, and is connected to the reservoir 16a via a suction pipe 201 which is driven by means of an automotive engine 200. The outlet of the pressure unit 16, through which the pressurized working fluid is fed, is connected to the inlet port 42i of the pressure control valve 18 via the supply line 52. A pressure regulating orifice 202 is disposed in the supply line 52 for suppressing pulsatile flow of the working fluid and whereby regulate the output pressure of the pressure unit 16 to be delivered to the pressure control valve 28. A feedback line 53 is connected to the upstream side of the pressure regulating orifice 202 at one end. The other end of the feedback line 53 is connected to the upstream side of the inlet of the pressure unit 16. Therefore, excessive fluid between the pressure unit 16 and the orifice 202 is fed back to the inlet side of the pressure unit. A pressure accumulator 203 is also connected to the supply line 52 to receive therefrom the pressurized fluid for accumulating the pressure. An one-way check valve 204 is disposed in the supply line 52 at the position upstream of the junction between the pressure accumulator 203 and the supply line 52. A pressure relief line 205 is also connected to the supply line 52 at the position intermediate between the pressure regulating orifice 202 and the one-way check valve 204, at one end. The other end of the pressure relief line 205 is connected to the drain line 54. A pressure relief valve 206 is disposed in the pressure relief line 205. The pressure relief valve 206 is responsive to the fluid pressure in the supply line 52 higher than a given value to drain part of the working fluid to the drain line for maintaining the pressure in the supply line 52 below the given pressure value. On the other hand, a shut-off valve 207 is disposed in the drain line 54. The shut-off valve 207 is also connected to the supply line 52 upstream of the one-way check valve 204 to receive therefrom the pressure in the supply line as a pilot pressure, via pilot line 208. The shut-off valve 207 is designed to be maintained at open position as long as the pilot pressure to be introduced through the pilot line 208 is held at a pressure level higher than or equal to a given pressure level. At the open position, the shut-off valve maintains fluid communication between the inlet side and outlet side thereof so that the working fluid in the drain line 54 may flow therethrough to the reservoir tank 16a. On the other hand, the shut-off valve 207 is responsive to the pilot pressure dropping below the given pressure level to be switched into shut-off position. At the shut-off position, the shut-off valve blocks fluid communication between the drain port 42o and the reservoir tank 16a. In parallel relationship to the shut-off valve, a pressure relief valve 209 is provided. The pressure relief valve 209 is disposed in a by-pass line 210 connecting the upstream side and downstream side of the shut-off valve 207. The pressure relief valve 209 is normally held at closed position to block fluid communication there through. On the other hand, the pressure relief valve 209 is responsive to a fluid pressure in the drain line 54 upstream thereof, higher than a set pressure, e.g. 30 kgf/cm², in order to establish fluid communication between the upstream side and downstream side of the shut-off valve to allow the excessive pressure at the upstream side drain line 54 to be drained therethrough. Therefore, the pressure relief valve 209 limits the maximum pressure at the set pressure. The set pressure of the pressure relief valve 209 corresponds to a predetermined offset pressure. An oil cooler 211 is disposed in the drain line 54 for cooling the working fluid returning to the reservoir tank 16a. Pressurized fluid supply operation to take place by the pressure source unit as set forth above will be discussed herebelow. While the automotive engine 200 is running, the fluid pump as the pressure unit 16 is driven. Therefore, the working fluid in the reservoir tank 16a is sucked via the suction pipe 201 and pressurized through the pressure unit 16. The pressurized working fluid is discharged from the outlet of the pressure unit 16 and fed to the pressure control valve 28 via the supply line 54 including the pressure regulating orifice 202 and the one-way check valve 204. When the pressure control valve 28 in a position of Fig. 2, the pressurized working fluid passes the pressure control valve and introduced into the working chamber 26d of the hydraulic cylinder 26. On the other hand, when the pressure control valve 28 is shifted to block communication between the supply line 52 and the working chamber 26d, the line pressure in the supply line increases. When the line pressure in the supply line 52 becomes higher than a set pressure of the pressure relief valve 206 in the pressure relief line 205, the excessive pressure higher than the set pressure is fed to the drain line 54 via the pressure relief valve 206 and thus returned to the reservoir tank 16a. The fluid pressure in the supply line 52 is also fed to the shut-off valve 207 via the pilot line 208. As set forth, the shut-off valve 207 is placed at open position as long as the pilot pressure introduced through the pilot line 208 is held higher than or equal to the set pressure thereof. Therefore, fluid communication between the pressure control valve 28 and the reservoir tank 16a is maintained. At this position, the working fluid is thus returned to the reservoir tank 16a via the drain line 54 via the shut-off valve 207 and the oil cooler 211. Since the shut-off valve 207, even at the open position, serves as a resistance to the fluid flow. Therefore, the fluid pressure in the drain line 54 upstream of the shut-off valve 207 becomes excessively higher, i.e. higher than the off-set pressure P₀. Then, the pressure relief valve 209 becomes active to open for allowing the excessive pressure of the working fluid to flow through the by-pass line 210. When the engine 200 stops, the pressure unit 16 ceases operation. By stopping of the pressure unit 16, the working fluid pressure in the supply line 52 drops. According to drop of the pressure in the supply line 52, the pilot pressure to be exerted to the shut-off valve 207 via the pilot line 208 drops. When the pressure in the pilot line 208 drops below or equal to the set pressure, the shut-off valve 207 is switched into shut-off position to block fluid communication therethrough. As a result, the fluid pressure in the drain line 54 upstream of the shut-off valve 207 becomes equal to the pressure in the working chamber 26d. Therefore, even when the working fluid leaks through a gap between the spool valve 48 and the inner periphery of the valve bore, it will not affect the fluid pressure in the working chamber 26d. This is advantageous to maintain the suspension characteristics of the suspension systems irrespective of the engine driving condition. As seen from Fig. 5, the working fluid pressure in the working chamber 26d of the hydraulic cylinder 26 varies according to variation of the current value of the control signal applied to the pressure control valve unit 28. As seen from Fig. 5, hydraulic pressure in the working chamber 26d varies between a maximum pressure P max which is saturation pressure of the pressure source unit 16 and a minimum pressure P min which is set at a magnitude in view of a noise component to be contained in the control signal. As seen from Fig. 5, the maximum hydraulic pressure P max corresponds to the maximum current value I max of the control signal and the minimum hydraulic pressure P min corresponds to the minimum current value I min of the control signal. Furthermore, the hydraulic pressure level as labeled P n represents neutral pressure at the neutral current I n . As seen, the neutral current value I n is set at an intermediate value between the maximum current value I max and the minimum current value I min . Fig. 6 shows an imaginary or theoritic block diagram of the control circuit for performing active suspension control in accordance with the present invention. As can seen from Fig. 6, the lateral acceleration sensor 23 is connected to the control unit 22. The control unit 22 performs anti-rolling suspension control principally based on the lateral acceleration indicative signal received from the lateral acceleration sensor 23. The control unit 22 includes a gain circuit 50f for amplifying the lateral acceleration indicative signal by a predetermined gain Kyf. The gain circuit 50f is directly connected to the front-left pressure control valve unit 28FL and connected to the front-right pressure control valve unit 28FR via an inverter 52f. Also, the control unit includes a gain circuit 50r for amplifying the lateral acceleration indicative signal of the lateral acceleration sensor 23 by a predetermined gain Kyr. The gain circuit 50r is directly connected to the rear-left pressure control valve 28RL and connected to the rear-right pressure control valve unit 28RR via an inverter 52r. Therefore, as can be appreciated herefrom, for the left and rear pressure control valve units, opposite polarity with equal magnitude of control signals are output for adjusting fluid pressure in working chambers of respectively associated hidraulic cylinders 26FL, 26FR, 26RL and 26RR. Since the control signals are to be derived for adjusting fluid pressure in the working chambers for suppressing vehicular rolling, the one of left and right hydraulic cylinders oriented at the inner side of the curve is softened and at the outer side of the curve is hardened. As seen from Fig. 7, signal values of the lateral acceleration indicative signal ga and gb are positive when lateral acceleration in response to right-hand steering is exerted and negative when lateral acceleration in response to left-hand steering is exerted. Absolute value of each of the lateral acceleration signals ga and gb varies in proportion to variation of the magnitude of lateral acceleration. On the other hand, respective vehicle height sensors 21 monitor relative distance between the suspension member 24 and the vehicular body 10 at respective relevant suspension systems 14. Fig. 8 shows practical implementation of the circuit construction of the preferred embodiment of the suspension control system employed in the shown embodiment of the actively controlled suspension system of Fig. 1 set forth above. The control unit 22 includes a microprocessor 100 as a main component of the control unit. The microprocessor 100 has per se well known construction including an input/output interface (I/O) 102, an arithmetic circuit 104 and a memory unit 106 mutually connected to each other via bus lines. The input/output interface 102 is connected to a plurality of analog-to-digital (A/D) converters 108A, 108B, 108C, 108D, 108E. The A/D converters 108A, 108B, 108C and 108D are respectively connected to the vehicle height sensors 21 of respective suspension systems 14FL, 14FR, 14RL and 14RR to receive therefrom analog form vehicular height indicative signals h₁, h₂, h₃ and h₄ and convert into digital signals having corresponding values. On the other hand, the A/D converter 108E is connected to the lateral acceleration sensor 23 for receiving the analog form lateral acceleration indicative signals ga and gb and converting into digital signals indicative thereof. The input/output interface 102 is also connected to a plurality of digital-to-analog (D/A) converters 110A, 110B, 110C and 110D which are, in turn, connected to driver circuits 112A, 112B, 112C and 112D. The D/A converters 110A, 110B, 110C and 110D receive digital form suspension control signals SC derived by the microprocessor 100 and output analog form control signals representative thereof. In practice, the analog form suspension control signals have current values variable depending upon the desired fluid pressure in the working chambers of respectively corresponding hydraulic cylinders 26. The driver circuits 112A, 112B, 112C and 112D are designed to convert the suspension control signals into analog form to output driver signals as current signals for driving the proportioning solenoid coils 60 at the corresponding magnitude. Here, the proportioning solenoid is generally connected to a driver circuit to receive driver current to drive a valve position for adjusting the pilot pressure at a desired pressure. The driver circuit adjusts a supply current toward a target current which is derived on the basis of a suspension control signal input from a control unit. Frequency characteristics of the supply current and the target current are so adjusted to have linear relationship. In such driver circuit, greater input gain is preferred for obtaining high response characteristics. On the other hand, for absorption of vibration energy, the hydraulic system in the pressure control valve preferably provides great input frequency characteristics versus input vibration. To achieve the high input frequency characteristics in absorbing the vibration energy, an orifice is provided in a path establishing communication between the outlet of the pressure control valve and a feedback chamber. This orifice tends to serve as resistance to the fluid flow and effective for providing high response in absorption of the vibration energy. However, on the other hand, this orifice serves as lag factor for response characteristics in attitude change suppressive mode operation in which the pilot pressure is controlled according to the driver current. As seen, the neutral current I n is set at an intermediate value between the maximum and minimum current values I max and I min . Fig. 9 shows the shown embodiment of the pressure control valve 28 as associated with the preferred embodiment of the driver circuit 112 employed for controlling the operation of the proportioning solenoid valve. The driver circuit 112 is connected to the solenoid coil 60 of the actuator assembly 29. The driver circuit 112 includes a phase advance circuit 151. The phase advance circuit 151 comprises a capacitor 152 and a resistor 153 arranged in parallel relationship to each other with respect to an input terminal P. The parallel circuit of the capacitor 152 and the resistor 153 is connected to a differential amplifier 154 and a parallel resistor 155. A series circuit of a resistor 156 and a capacitor 157 is connected between the differential amplifier 154 and the junction of the parallel circuit of the capacitor 152 and the resistor 153. This series circuit is connected to the ground. The output of the phase advance circuit 151 is connected to a differential amplifier 159 via a resistor 158. The differential amplifier 159 has the other input terminal connected to a junction between resistors 160 and 161. The output of the differential amplifier 159 is connected to the base electrode of a transistor 163 via a resistor 162. The emitter electrode of the transistor 163 is connected to a junction between the resistor 161 and a resistor 164 which is connected to the ground at the other end. The collector electrode of the transistor 163 is connected to a power source battery 165 via the solenoid coil 60. The driver circuit 112 set forth above operates in the following manner. In response to vehicular body attitude change indicative input, the control unit 22 outputs the suspension control signal. The control signal is supplied in a form of a current signal having a current value representative of desired magnitude of the pilot pressure. The current value of the control signal will be hereafter referred to as a ''target current I₁''. The target current I₁ is supplied to the non-inverting input terminal of the differential amplifier 154 of the phase advance circuit 151 through the input terminal P and the parallel circuit of the capacitor 152 and the resistor 153. Output of the differential amplifier 154 is fed back through the feedback resistor 155. The fed back output is summed with the input current at the summing junction connected to the non-inverting input terminal. As a result, phase of the output I₃ of the differential amplifier 154 is shifted in advancing direction as shown in Fig. 11. The phase advanced output current I₃ of the differential amplifier 154 is fed to the non-inverting input terminal of the differential amplifier 159. The output of the differential amplifier 159 is supplied to the base electrode of the transistor 163. This causes rising of the potential at the base electrode across the potential at the emitter electrode to cause the transistor turning ON. Therefore, the current corresponding to the current value I₃ is applied to the solenoid coil 60 to energize the latter. Therefore, the solenoid coil 60 is energized by the current with a primary lead phase. Here, the current I₂ actually flowing through the solenoid coil can be detected as a terminal voltage at the resistor 164. Therefore, by connecting the resistor 164 to the inverting input terminal of the differential amplifier 159 via the resistor 161, the current values I₃ and I₂ can be compared and adjusted toward the current value I₃. With such construction of the driver circuits 112 in the shown embodiment, satisfactorily high response characteristics can be obtained which will be further discussed in order to assist clear understanding of the invention. Fig. 10 shows a model of the vehicle in which the front-left and front-right wheels are represented by a front wheel oriented on the longitudinal axis and the rear-left and rear-right wheels are represented by a rear wheel oriented on the longitudinal axis, for simplification of the following discussion. As seen, the lateral acceleration sensor 23 is provided at a position on the longitudinal axis and shifted frontwardly in a magnitude of e from the gravity center y. In the model of Fig. 10, the following condition is set for analysis: Vehicular Weight: M; Yawing Moment at about Gravity Center: I; Wheel Base: ℓ; Distance Between Front wheel and Gravity Center: a; Distance Between Rear wheel and Gravity Center: b; Vehicle Speed: V; Cornering Power at Front Wheel: C₁; Cornering Power at Rear Wheel: C₂; Steering Angle at Front Wheel: ϑ; Lateral Displacement of Gravity Center: y Here, fundamental equations are: M (α + V) = f₁ + f₂ (1) Iα = af₁ + bf₂ (2) where α + V is lateral acceleration at the gravity center,  is yawing rate, α is yawing angular acceleration, and f₁ and f₂ are cornering force at the front and rear wheels. The cornering forces f₁ and f₂ at the front and rear wheels can be illustrated as follows: f₁ = C₁ {ϑ - (a + y)/V} (3)a f₂ = C₂ { - (α + b)/V} (3)b wherein ϑ - (a + y)/V is lateral slip angle at the front wheel and - (α + b)/V is lateral slip angle at the rear wheel. Here, the lateral slip angle β at the gravity center can be illustrated by: β = y/V The lateral slip angle β increases in counterclockwise direction. For deriving the lateral acceleration α + V and yawing rate  in relation to the actual steering angle ϑ, Laplace transformation is made (α + V)/ϑ = {A₂ω n ² (sd² + 2 ε₂ω₂s + ω₂²/ω₂²}/(s² + 2 ε n ω n s + ω n ²) (4) /ϑ =A₁ω n ²(1 + τ₁s)}/(s² + 2 ε n ω n s + ω n 2) (5) where ω n = (ℓ/V) C₁C₂(1 + KsV²)/IM 1/2 ε n = {(C₁+C₂)I + (a²C₁+b²C₂)/M}/2ℓ IMC₁C₂(1+KsV²) 1/2 A₂ = V²/{ℓ(1 + KsV²)} A₁ = V/{ℓ(1 + KsV²)} ω₂ = (ℓC₂/I) 1/2 ε₂ = (b/2V)(ℓC₂/I) 1/2 τ₁ = (aMV)/ℓC₂ Ks = (M/ℓ²)(b/C₁ - a/C₂) When the lateral acceleration sensor 23 is facilitated at the orientation as shown in Fig. 10, the lateral acceleration monitored by the lateral acceleration sensor 23 becomes a sum of the lateral acceleration α + V and lateral acceleration eα generated by yawing angular acceleration α. Therefore, the lateral acceleration (α + V + eα) . Therefore, when transferring function Gs is set at a ratio of the lateral acceleration (α + V) at the gravity center and the lateral acceleration (α + V + eα) monitored by the lateral acceleration sensor, the transferring function can be described by: Utilizing the equation (6) and with taking vehicle speed V and the distance e between the gravity center and the sensor position as parameters, characteristics of gain and phase as illustrated in Figs. 11(a), 11(b) and 12(a), 12(b) could be obtained. As seen from Figs. 11(a), 11(b) and 12(a), 12(b), gain became greater and phase advance was improved by increasing the distance e. Phase advance is also improved at higher vehicle speed for quick steering operation which has general characteristics of 1 to 2 Hz. On the other hand, rolling rate characteristics γ in relation to the actual steering angle ϑ is variable depending upon steering frequency and the distance e, as shown in Figs. 13(a) and 13(b). Except for the racing cars, usual steering frequency is less than or equal to 2 Hz. In order to make γ/ϑ smaller, the distance e has to be selected at an appropriate distance. The appropriate distance in this view point is variable depending upon the specification of the vehicles to apply. However, in general, for the passenger's car, the frequency at around 1 Hz is important. As will be appreciated, the roll gain will become excessive when the distance e becomes excessively longer. Therefore, the preferred range of the distance e is 20 to 40 cm from the gravity center. 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.

Claims: An actively controlled suspension system for an automotive vehicle, comprising: a cylinder (26) disposed between a vehicle body (10) and a suspension member (24) which rotatably supports a road wheel, said cylinder defining a variable pressure working chamber filled with a pressure medium for generating a damping force resisting relative displacement between said vehicle body and said suspension member, a pressure of said pressure medium being variable between a predetermined maximum value and a predetermined minimum value across a predetermined neutral value, said cylinder being provided for each of the suspension systems respectively associated with the front and rear wheels (11); a pressure source circuit (16,35,37) connected to said working chamber for supplying pressure medium and including a pressure source feeding pressurized medium through said circuit; a pressure control valve (28) disposed between said pressure source and said working chamber and having a valve with a valve position variable between a first mode for increasing pressure of said pressure medium within said working chamber, a second mode for decreasing pressure of said pressure medium within said working chamber and a third mode for maintaining pressure of said pressure medium constant, said pressure control valve being provided for each of said suspension systems respectively associated with said front and rear wheels; a lateral acceleration sensor (23) for monitoring a lateral acceleration exerted on the vehicle body to cause vehicular attitude change and produce a sensor signal indicative thereof; and a controller means (22) for receiving said sensor signal for deriving a suspension control signal for operating said pressure control valve at one of said first, second, or third mode positions for regulating vehicular attitude based on said sensor signal; characterized in that said lateral acceleration sensor is oriented on the vehicular body at a position frontwardly offset from a center of gravity of the vehicle body to provide the sensor signal which includes a component created by yawing angular acceleration acting on the vehicle body at the position of the acceleration sensor as a function of the distance of said offset position relative to said center of gravity. An actively controlled suspension system as claimed in claim 1, characterized in that the frontward offset magnitude of said lateral acceleration sensor is derived so as to advance a phase of the control system by a preselected magnitude. An actively controlled suspension system as claimed in claim 2, characterized in that said frontward offset magnitude of said lateral acceleration sensor relative to the gravity center is derived so that the lateral acceleration sensor is additionally subject to a yawing force in the magnitude advancing a phase of the control system by the preselected magnitude.

Inventors: AKATSU, YOHSUKE, FUJIMURA, ITARU 2-2, KOYAMA-SO, 220, YATSU-MACHI, FUKUNAGA, YUKIO B-50, HIGASHIZUSHI-HAITSU, 8, FUKUSHIMA, NAOTO, SATO, MASAHARU
Applicants: NISSAN MOTOR CO., LTD.
Application year: 1989
Main classification: B60G17
IPC codes: B60G17