ACTIVELY CONTROLLED AUTOMOTIVE SUSPENSION SYSTEM WITH IMPROVED CORNERING CHARACTERISTICS

Abstract: A suspension system for an automotive vehicle employs an active suspension control for controlling the suspension characteristics according to the vehicle driving condition. Active suspension control is performed by adjusting offset of a gravity center for adjusting load distribution at respective vehicular wheels. The load distribution is adjusted so that the suspension characteristics can be changed to over-steer characteristics when the vehicle speed demand is higher than a given level so as to allow drift of the vehicle at a corner. In order to assure driving stability, the suspension system is further adjusted to suppress roll during drift of the vehicle.

Description: The present invention relates to a suspension system for an automotive vehicle comprising: suspension units disposed between a vehicle body and a plurality of suspension members supporting respectively associated road wheels, each of said suspension units including actuator means adjustable by means of hydraulic fluid; a hydraulic circuit connected to said actuator means for supplying pressurized working fluid thereto; sensor means for monitoring predetermined vehicle driving torque indicative parameters to produce a driving torque indicative signal and sensor means for monitoring lateral vehicle acceleration to produce a lateral acceleration Indicative signal; and controller outputting front and rear actuator target stroke signals as a function of said driving force and lateral acceleration indicative signals to cause adjustment of actuator position causing a change in load distribution between said suspension units associated with each road wheel so as to actively oppose vehicle rolling during cornering. Generally, a typical construction of a suspension assembly comprises a suspension coil spring 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-stabilizer for suppressing vehicular rolling motion. Some of the many kinds of automotive suspension systems include hydraulic circuits associated with fluid chambers in hydraulic shock absorbers for controlling the 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 chamber 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. WO-A-8 402 886 discloses a suspension system having suspension units disposed between a vehicle body and a plurality of suspension members supporting road wheels. Each suspension unit includes an actuator means adjustable by means of hydraulic fluid. A hydraulic circuit is connected to said actuator means for supplying pressurized working fluid thereto. Sensors monitor predetermined vehicle driving conditions and produce driving condition signals from which a controller derives a signal to cause an adjustment of the actuator positions to change the load distribution between said suspension units. The controller produces pressure control signals for each suspension unit to obtain a desired vehicle attitude and suspension unit load distribution for optimum vehicle driving characteristics. On the other hand, as is well known, driving torque to rotate the road wheel exerts to the tread of the road wheel as longitudinal force. In cornering, lateral force acts on the tread of the road wheel due to centrifugal force. The combined force of the driving torque and the lateral force determines the cornering characteristics of the vehicle. Namely, cornering force as the counterforce against the combined force as set forth is generated. Generally, the magnitude of this counterforce to be generated is determined by road/tyre friction. The magnitude of the road/tyre friction is variable depending upon the load applied to the road wheel. Therefore, combination of the combined force and the counterforce thereagainst at each vehicular wheel will determine the cornering characteristics of the vehicle. In other words, the maximum counterforce to be generated at each wheel is determined by road/tyre friction. The distribution of the friction magnitude at respective road wheel is variable depending upon the shift of a gravity center during cornering due resulting change of load distribution. When the lateral force becomes greater than the maximum lateral component of the counterforce, the tyre causes break of traction to cause lateral slide to increase over-steer characteristics. On the other hand, as long as the road/tyre traction is maintained, under-steer characteristics may increase according to increasing of driving torque as the longitudinal force. This prevents the vehicle to pass the corner at high speed. Therefore, in order to optimize cornering characteristics of the vehicle at relatively high speed, balance of the combined force of the longitudinal and lateral forces and the counterforce to cause slight lateral slid Is required. However, in order to obtain such optimum cornering characteristics, quite high driving technic, such as delicate accelerator pedal control and delicate steering operation synchronized with accelerator pedal control is required. It is an object of the present invention to provide an actively controlled automotive suspension system which automatically controls the suspension characteristics for obtaining optimum cornering characteristics to allow driving of the vehicle through the corner at relatively high speed. In order to accomplish the aforementioned object of the invention, a suspension system for an automotive vehicle according to the present invention is characterized in that said controller increases the magnitude of load shift between the rear wheels as compared to that between the front wheels as the driving force indicative signal increases such that at a predetermined driving force indicative value the magnitude of load shift between the rear wheels exceeds that between the front wheels thereby changing the vehicular steering characteristics to over-steer to enable the vehicle to actively tail slide during cornering. The system according to the invention employs an active suspension control for actively controlling the suspension characteristics according to the vehicle cornering conditions. Active suspension control is performed by adjusting the load shift between the rear wheels, compared to that between the front wheels. Thus the cornering characteristic is changed to over-steer when the vehicle speed demand is higher than a given level so as to allow a drift of the vehicle at a corner. In the preferred construction, in order to assure driving stability, the suspension system is further adjusted to suppress vehicular rolling during drift of the vehicle. In a preferred construction, the hydraulic means comprises a hydraulic cylinder disposed therein a reciprocating piston which defines a fluid chamber within the hydraulic cylinder and responsive to the hydraulic pressure within the fluid chamber, and a valve means interposed between the hydraulic circuit and the fluid chamber for supplying and draining pressurized working fluid to and from the fluid chamber for adjusting fluid pressure in the fluid chamber according to the pressure control signal. The pressure control valve includes a valve body and an electrically operable actuator for shifting the valve body for adjusting fluid pressure in the fluid chamber according to the pressure control signal. The sensor means may include means for monitoring lateral force exerting on the vehicle for producing a lateral force indicative signal and means for monitoring vehicular speed demand indicative parameter to produce a vehicular speed indicative sensor signal. The controller is responsive to the lateral force indicative signal and vehicular speed demand indicative sensor signal to derive the load distribution. The controller is responsive to vehicular speed demand indicative parameter having a value greater than a predetermined value to change vehicular steering characteristics to over-steer characteristics. The DE-A-3 512 047 discloses a suspension system whereby in order to achieve desired driving characteristics, load distribution is adjusted between the front road wheels and the rear road wheels. This is achieved by separately adjusting the height of the vehicle body with respect to the wheels. A greater height at the rear axle results in over-steer characteristics. In the following, embodiments of the invention will be explained with reference to 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 section of one example of a hydraulic suspension assembly in combination with a hydraulic pressure control valve associated with the suspension assembly for adjusting the latter; Fig. 3 is a block diagram of the suspension control system in the actively controlled automotive suspension system of Fig. 1; 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 suspension and Fig. 4(B) shows the valve position hardening; 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 an explanatory illustration showing the forces to be exerted on a road wheel during cornering; Fig. 7 is a graph showing variation of cornering force according to shifting of the gravity center or variation of the load applied to road wheels; and Fig. 8 is an illustration of variation of attitude of vehicle during cornering at high speed. Referring now to the drawings, 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 comprises a suspension member 13, such as a suspension link, a suspension arm and so forth, and a suspension assembly 15 which is incorporated between the vehicle body 12 and the suspension member 13. The suspension assembly 15 has a hydraulic cylinder 15A which serves as an actuator, a coil spring 16. The hydraulic cylinder 15A has a hollow cylinder housing 15a filled with a viscous working fluid and a piston 15b 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 15c extends through one end of the cylinder housing 15a. The other end of the cylinder housing 15a is connected to one of the vehicle body 12 and the suspension member 13. On the other hand, the end of the piston rod 15c is connected to the other of the vehicle body 12 and the suspension member 13. 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. The hydraulic circuit 19 includes a supply line 19s and a drain line 19d. High pressure accumulators 22f and 22r are connected to the supply line 19s and low pressure accumulators 23f and 23r are connected to the drain line 19d. 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. The controller 30 is connected to stroke sensors 17 which are respectively interposed between the vehicle body 12 and the corresponding suspension members 13. Each stroke sensor comprises a potentiometer, for example, and is designed to monitor the relative distance between the vehicle body 12 and the associated suspension member 13 to produce a stroke indicative signal. The stroke indicative signal is variable of the signal value to indicate magnitude of relative displacement of the vehicle body 12 and the suspension member 13 from the neutral position. Namely, when the relative distance expands, the stroke indicative signal value becomes positive and increases as increasing of the expanded magnitude, and when the relative distance decreased from the neutral position, the stroke indicative signal value decreases as decreasing of the relative distance. The controller 30 is also connected to a lateral acceleration sensor 24, a transmission gear position sensor 25 and a throttle angle sensor 26. The lateral acceleration sensor 24 monitors acceleration laterally exerted on the vehicle to produce a lateral acceleration indicative signal α. The lateral acceleration indicative signal value α has a voltage variable depending upon the magnitude and direction of the lateral acceleration. In the shown embodiment, the positive voltage of the lateral acceleration indicative signal increases as increasing the leftward lateral acceleration, and the negative voltage decreases as increasing the rightward lateral acceleration increases. The transmission gear position sensor 25 is associated with an automotive power transmission to detect the shift position of the transmission to produce a transmission gear position indicative signal g. The transmission gear position indicative signal has a voltage value depending upon the transmission gear position. On the other hand, the throttle angle sensor 26 monitors the angular position of a throttle valve (not shown) linked with an accelerator pedal (not shown) in per se well known manner, to produce a throttle angle indicative signal ϑ. The transmission gear position sensor 25 and the throttle angle sensor 26 constitute a vehicle driving condition indicative parameter detecting means. On the other hand, the lateral acceleration sensor 24 constitutes a vehicle running condition indicative parameter detecting means. The controller 30 receives the aforementioned stroke indicative signals from the stroke sensors 17, a lateral acceleration indicative signal ⊖ from the lateral acceleration sensor 24, the transmission gear position indicative signal g from the transmission gear position sensor 25 and the throttle angle indicative signal ϑ of the throttle angle sensor 26 to derive control signal based thereon. The controller 30 outputs the control signal to the pressure control valves 18. 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 forementioned outlet port 18d, an inlet port 18b and a drain port 18c. Respective inlet port 18b, the drain port 18c and the outlet port 18d are connected to a valve bore 18a defined within the valve housing 18A. A valve spool 19 is disposed within the valve bore 18a for thrusting movement therein. The valve spool 19 has first, second and third lands 19a, 19b and 19c. As will be seen from Fig. 2, the third land 19c has smaller diameter than that of the first and second lands 19a and 19b. The third land 19c defines a fifth pressure control chamber 18h which is connected to the drain port 18c via a drain path 18f. An actuation rod 29 is also disposed within the valve bore 18a. The actuation rod 20 opposes the second land 19b in spaced apart relationship to define a second pressure control chamber 18i which is connected to the drain port 18c via a drain path 18e. An annular pressure chamber 18j is defined between the first and second lands 19a and 19b. The pressure chamber 18j is constantly 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. On the other hand, an annular 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 21 is interposed between the actuation rod 20 and the valve spool 19. The actuation rod 20 contacts with an actuator rod 22a of an electrically operable actuator 22 which comprises an electromagnetic solenoid. The solenoid 22 comprises a proportional solenoid. Fig. 3 shows the controller 30 in a form of a block diagram. The controller 30 generally comprises a first driving parameter indicative value generator 31, a second driving parameter indicative gain constant generating circuit 32, multipliers 33, 34f and 34r, inverters 35f and 35r, comparators 36FL, 36FR, 36RL and 36RR. The first driving parameter indicative value generator circujit 31 derives a first driving parameter indicative value. In the shown embodiment, the first driving parameter indicative value generator circuit 31 receives the transmission gear position indicative signal g from the transmission gear position sensor 25, The first driving parameter indicative value generator 31 derives a reciprocal or the gear ratio corresponding to the transmission gear position as represented by the transmission gear position indicative signal. In practice, the first driving parameter indicative value generator comprises a plurality of value generators, each of which is designed to produce a value corresponding to the reciprocal of the gear ratio at the corresponding transmission gear position. In response to the transmission gear position indicative signal g from the transmission gear position sensor 25, one of the value generator which corresponds to one of the transmission gear position as indicated in the transmission gear position indicative signal, is triggered to output a first driving value indicative signal N. The first driving coefficient indicative signal N is fed to the multiplier 33. The multiplier 33 also receives the throttle angle indicative signal T from the throttle angle sensor 26. The multiplier 33 multiplies the first value indicative signal N with the throttle angle indicative signal Θ to derive an assumed driving force indicative value T. The assumed driving force indicative value T is fed to the driving parameter indicative gain constant generating circuit 32. The driving parameter indicative gain constant generating circuit 32 derives a gain constant based on the second driving parameter indicative value T. In order to derive the gain constant, the driving parameter indicative gain constant generating circuit 32 has a gain constant data table which is in a form of look-up table. The gain constant data table is looked up in terms of the assumed driving force T. As will be seen in Fig. 3, the gain constant K₁ for the front wheels 14FL and 14FR gradually decreases according to increasing of the assumed driving force T. The gain constant K₁ for the front wheels 14FL and 14FR is hereafter referred to as ''front gain constant''. On the other hand, the gain constant K₂ for the rear wheels 14RL and 14RR increases according to increasing of the assumed driving force T. The gain constant K₂ for the rear wheels 14RL and 14RR is hereafter referred to as ''rear gain constant''. When the assumed driving force T is zero, the front gain constant K₁ is greater than the rear gain constant K₂. The difference of the front and rear gain constants K₁ and K₂ decreases as increasing of the assumed driving force. The rear gain constant K₂ increases across the front gain constant K₁ at a given assumed driving force. The driving parameter indicative gain constant generating circuit 32 outputs a front gain indicative signal representative of the front gain constant to the multiplier 34f and a rear gain indicative signal representative of the rear gain constant to the multiplier 34r. On the other hand, the multipliers 34f and 34r are connected to the lateral acceleration sensor 24 to receive therefrom a lateral acceleration indicative signal having a lateral acceleration indicative value α. The lateral acceleration indicative value α of the lateral acceleration indicative signal is multiplied by the front and rear gain constants K₁ and K₂ respectively in the multipliers 34f and 34r. The multiplier 34f thus derives a magnitude of a shift of load between the front wheels 14FL and 14FR on the basis of the front gain constant K₁ as a front driving condition indicative variable, and the lateral acceleration. The value derived by the multiplier 34f thus indicative of the magnitude of shifting of the load at respective front-left and front-right wheels and, in turn, indicative of the magnitude of variation of the relative distance between the vehicle body and the suspension member. Therefore, the multiplier outputs a front target stroke signal Wf indicative of the derived magnitude of the load shift between the front wheels. On the other hand, the multiplier 34r derives a magnitude of a shift of load between rear wheels 14RL and 14RR. The value derived by the multiplier 34r thus indicative of the magnitude of shifting of the load at respective rear-left and rear-right wheels and, in turn, indicative of the magnitude of variation of the relative distance between the vehicle body and the suspension member. Therefore, the multiplier outputs a rear target strobe signal Wr having a value representative of the rear load shifting magnitude. The multiplier 34f feeds the front load shift control signal Wf to a comparator 36FL for the front-left wheel and to a comparator 36FR via an inverter 35f. Therefore, the polarity of the front load shift control signals Wf to be applied to the comparators 36FL and 36FR becomes alternative. Similarly, the multiplier 34r feeds the rear load shift control signal Wr to a comparator 36RL for the front-left wheel and to a comparator 36RR via an inverter 35r. Therefore, the polarity of the front load shift control signals Wr to be applied to the comparators 36RL and 36RR becomes alternative. Each comparators 36FL, 36FR, 36RL and 36RR is connected to the stroke sensor 17 to receive the stroke indicative signal DS FL , DS FR , DS RL and DS RR . The comparators 36FL, 36FR, 36RL and 36RR respectively compare the control signals Wf, -Wf, Wr and -Wr with respective stroke indicative signals DS FL , DS FR , DS RL and DS RR in order to derive error indicative signals ε FL , ε FR , ε RL and ε RR . The error indicative signals ε FL , ε FR , ε RL and ε RR respectively representative of the magnitude of variation of the stroke of the hydraulic cylinders 15 at respective front-left, front-right, rear-left and rear-right wheels 14FL, 14FR, 14RL and 14RR. The comparators 36FL, 36FR, 36RL and 36RR feeds the error indicative signals ε FL , ε FR , ε RL and ε RR to respectively associated driver circuits 37FL, 37FR, 37RL and 37RR. The driver circuits produce drive signals for respectively associated actuators 22 of the pressure control valves 18. The actuators 22 are thus actuated at the magnitudes respectively corresponding to the error values indicated by the error indicative signal ε FL , ε FR , ε RL and ε RR , in order to reduce the error to zero. In the operation of the aforementioned controller 30, the input from the lateral acceleration sensor 24 is maintained zero as long as the vehicle travels in substantially straight and thus substantially no lateral force is exerted on the vehicle. As a result, the output as the front and rear target stroke indicative signals, of the multipliers 34f and 34r becomes substantially zero. At this time, since the relative distances between the vehicle body and the suspension members are maintained substantially at the neutral position, since no lateral force is applied. Therefore, the values of the stroke indicative signals DS FL , DS FR , DS RL and DS RR to be inputted to the comparators 36FL, 36FR, 36RL and 36RR from respectively corresponding stroke sensors 17 are held substantially zero. Therefore, the output value of the comparators 36FL, 36FR, 36RL and 36RR are held zero. Therefore, the error indicative signals ε FL , ε FR , ε RL and ε RR output from the comparators 36FL, 36FR, 36PL and 36RR are held substantially zero. As a result, the driver circuit 37FL, 37FR, 37RL and 37RR are held inoperative to maintain the actuators 22 of respective pressure control valves 18 in place. On, the other hand, when a vehicular steering operation is performed to change driving direction, lateral force is exerted on the vehicle body to generate the lateral acceleration. The lateral acceleration sensor 24 then outputs the lateral acceleration indicative signal with a value α. As a result, the output values Wf and Wr as the front and rear target stroke indicative values become (α x K₁) and (α x K₂) respectively. At the same time, steering operation may cause vehicular rolling to cause variation of the relative distances between the vehicle body and suspension members at the positions of the front-left, front-right, rear-left and rear right wheels. Therefore, the stroke indicative signal values varies from zero. For instance, if left-hand steering is performed through a steering wheel of the vehicle, the rightward centrifugal force is exerted on the vehicle to causing vehicular rolling to lower the right-side of the vehicle body and to rise the left-side of the vehicle body. Therefore, the stroke indicative signal values from the stroke sensors 17 associated with the front-left and rear-left wheels become positive and the stroke indicative signal values from the stroke sensors 17 associated with the front-right and rear-right wheels become negative. On the other hand, as set forth above, the lateral acceleration indicative signal values from the lateral acceleration sensor 24 becomes negative value (-α). Therefore, the front and rear target strokes Wf and Wr become negative. The negative value of the front traget stroke (-Wf) is input to the comparator 36FL to be compared with the positive value of the stroke indicative signal DS FL . Therefore, the error indicative signal ε FL indicative of the diference of the front target stroke indicative signal and the stroke indicative signal to reduce magnitude of rising but rather cause downward shift down across the neutral position at the front-left wheel toward the front target stroke (-Wf). On the other hand, the inverted front traget stroke (+Wf) is input to the comparator 36FR. The comparator 36FR also receives the negative value of the stroke indicative signal from the stroke sensor 17 associated with the front-right wheel. As a resuly, the error indicative signal ε FR will be derived to adjust the relative distance between the vehicle body and the front-right suspension member toward the inverted front target stroke (+Wf). Similarly, The negative value of the rear traget stroke (-Wr) is input to the comparator 36RL to be compared with the positive value of the stroke indicative signal DS RL . Therefore, the error indicative signal ε RL indicative of the difference of the rear target stroke indicative signal and the stroke indicative signal to reduce magnitude of rising but rather cause downward shift down across the neutral position at the front-left wheel toward the rear target stroke (-Wr). On the other hand, the inverted rear traget stroke (+Wr) is input to the comparator 36FR. The comparator 36FR also receives the negative value of the stroke indicative signal from the stroke senior 17 associated with the rear-right wheel. As a result the error indicative signal ε RR will be derived to adjust the relative distance between the vehicle body and the REAR-right suspension member toward the inverted rear target stroke (+Wr). Based on the error indicative signals ε FL , ε FR , ε RL , and ε RR as set forth above, the driver circuits 37FL, 37FR, 36RL and 37RR outputs drive signals to the respectively associated actuator 22. The driver signal value is variable to control the position of the valve spool 19. The pressure control valve 18 is designed to supply the working fluid pressure P ₀ (shown in Fig. 5) to the upper fluid chamber 15d. As will be seen from Fig. 5, the working fluid pressure to be supplied to the upper fluid chamber 15d increases as increasing the drive signal voltage V and decreases as decreasing the drive signal voltage. By increasing the working fluid pressure in the upper fluid chamber 15d of the hydraulic cylinder 15a, the piston 15 strokes expand the relative distance between the vehicle body and the suspension member, which piston stroke will be hereafter; referred to as ''expansion stroke''. On the other hand, when the working fluid pressure in the upper fluid chamber 15d is decreased, the piston strokes to reduce the relative distance between the vehicle body and the associated suspension member. 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. 2, 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. On the other hand, as will be seen from Figs. 6 and 7, driving torque T to rotate the road wheel exerts to the tread of the road wheel as longitudinal force. On the other hand, in cornering, lateral force acts on the tread of the road wheel due to centrifugal force. The combined force F of the driving torque and the lateral force determines the cornering characteristics or the vehicle. Namely, cornering force F which is combined force of a lateral counter force S and the driving torque T, as the counter force against the combined force of the centrifugal force and the driving torque as set forth is generated. Generally, the magnitude of this counter force F to be generated is determined by road/tyre friction, magnitude of which is illustrated by a circle in Fig. 6 for the purpose of illustration. The magnitude of the road/tyre friction is variable depending upon the load W applied to the road wheel. Therfore, combination of the combined force and the counter force thereagainst at each vehicular wheel will determine the cornering characteristics of the vehicle. In other words, the maximum counter force Fmax to be generated at each wheel is determined by road/tyre friction. Here, assuming the loads at left and right wheels are respectively W L and W R , the average load of the left and right wheel loads W L , and W R is W O and a magnitude of shift of the load between the left and right wheels with respect to the center thereof is ΔW. the maximum counter force Fmax decreases according to increasing of the load shift magnitude ΔW. In other words, the distribution of the friction magnitude at respective road wheel is variable depending upon the shift of a gravity center during cornering due resulting change of load distribution. When the lateral force becomes greater than the maximum lateral component of the counter force, the tyre causes break of traction to cause lateral slide to increase over-steer characteristics. On the other hand, as long as the road/tyre traction is maintained, under-steer characteristics may increase according to increasing of driving torque as the longitudinal force. This prevents the vehicle to pass the corner at high speed. Therefore, in order to optimize cornering characteristics of the vehicle at relatively high speed, balance of the combined force or the longitudinal and lateral forces and the counter force to cause slight lateral slide is required. However, in order to obtaine such optimum cornering characteristics, quite high driving technic, such as delicate accelerator pedal control and delicate steering operation synchronized with, accelerator pedal control is required. Namely, assuming the load shift magnitude increase from ΔW (W L1 , W R1 ) to ΔW (W L2 , W R ), the maximum counter force varies from ΔF₁ to ΔF₂. This means tendency of over-steer in cornering increases by shifting load between left and right wheels. By the control operation of the controller 30, the rolling direction of the vehicle is reversed from the normal rolling direction. This changes the steering characteristics of the vehicle to be inclined toward over-steer characteristics. Therefore, as shown in Fig. 8, when the vehicle travels along a corner with relatively high driving force, the vehicular suspension characteristics changes toward the over-steer characteristics to cause tail-slide to drift without requiring any attention and/or skill of the driver. Therefore, the present invention fulfills all of the objects and advantages sought therefore.

Claims: A suspension system for an automotive vehicle comprising: suspension units (11FL,11FR,11RL,11RR) disposed between a vehicle body (12) and a plurality of suspension members (13) supporting respectively associated road wheels (14), each of said suspension units (11) including actuator means (15A) adjustable by means of hydraulic fluid; a hydraulic circuit (18,19s,19d, 20,21) connected to said actuator means for supplying pressurized working fluid thereto; sensor means (25,26) for monitoring predetermined vehicle driving torque indicative parameters (g,ϑ) to produce a driving torque indicative signal (T) and sensor means (24) for monitoring lateral vehicle acceleration to produce a lateral acceleration indicative signal (α); and controller outputting front and rear actuator target stroke signals (Wf,Wr) as a function of said driving force and lateral acceleration indicative signals to cause adjustment of actuator position causing a change in load distribution between said suspension units associated with each road wheel so as to actively oppose vehicle rolling during cornering, characterized in that said controller (30) increases the magnitude of load shift (ΔW) between the rear wheels as compared to that between the front wheels as the driving force indicative signal increases such that at a predetermined driving force indicative value the magnitude of load shift between the rear wheels exceeds that between the front wheels thereby changing the vehicular steering characteristics to over-steer to enable the vehicle to actively tail slide during cornering. A suspension system as claimed in claim 1, characterized in that said actuator means includes a hydraulic cylinder (15A) having disposed therein a reciprocating piston (15b) which defines a fluid chamber within said hydraulic cylinder and responses to the hydraulic pressure within said fluid chamber, and a valve means (18) interposed between said hydraulic circuit (18,19s, 19d,20,21) and said fluid chambers for supplying and draining pressurized working fluid to and from said fluid chamber for adjusting fluid pressure in said fluid chamber according to said actuator target signals. A suspension system as claimed in claim 2, characterized in that said pressure control valve (18) includes a valve body (19a,19b,20) and an electrically operable actuator for shifting said valve body for adjusting fluid pressure in said fluid chamber according to said actuator target signal.

Inventors: FUKUSHIMA, NAOTO, TORII, SHUUJI
Applicants: NISSAN MOTOR CO., LTD.
Application year: 1987
Main classification: B60G17
IPC codes: B60G17