Suspension control system for automotive vehicle with adjustment of wheel slippage dependent wheel load distribution

Abstract: A suspension control system includes means for monitoring wheel slippage on each of individual wheels as a parameter representative of road/tire traction. Suspension control is performed generally according to the vehicle driving conditions reflected in various suspension control parameters, such as vehicular speed, lateral acceleration and so forth in order to adjust spring coefficient or damping characteristics of a suspension system which is provided between a vehicle body and a suspension member supporting each individual wheel in order to adjust wheel load distribution and whereby to minimize variation of vehicular attitude change. Suspension control is further performed with taking the road/tire friction as one of suspension control parameters in order to provide traction dependent wheel load distribution.

Description: The present invention relates to a suspension control system according to the preamble of claim 1. A suspension control system of such a kind is disclosed by EP-A 0 231 025. Japanese Patent First (unexamined) Publication (Tokkai) Showa 62-275814 discloses a sort of an active suspension system which acts in active manner for suppressing vehicular body attitude change according to vehicle driving condition and thus achieves both of vehicular riding comfort and driving stability. Such prior proposed suspension control system adjusts wheel load distribution in order to suppress vehicular body attitude change, such as pitching and rolling. Particularly, the disclosed invention in the prior proposal concerns anti-rolling suspension control for adjusting wheel distribution for left and right side wheels so as to appropriately adjusting cornering force. Namely, according to the prior proposal, wheel load distribution is so adjusted as to resist against yawing moment which is induced by lateral acceleration generated by driving and braking force exerted on the vehicle. Such prior proposed system is successful in accomplishing minimization of variation of cornering force and whereby providing better cornering performance of the vehicle. On the other hand, such prior proposed still holds a drawback to be improved for providing better balance or higher stability of the vehicle. Namely, in the prior proposal, road surface friction or road/tire traction which reflects wheel slippage at each individual wheel, has not been taken into account in controlling the wheel load distribution. The factor of road/tire traction becomes particularly important for providing satisfactorily high vehicular driving stability during cornering or braking at a condition close to critical points. Namely, when traction Is relatively low, there is a possibility of breaking of traction for causing abrupt vehicular attitude change for falling the vehicle in unstable condition. The above mentioned EP-A 0 231 025 discloses a vehicle attitude control device in which anti-skid control is carried out during braking, and acceleration slip control is carried out during acceleration of the vehicle. Said anti-skid control is performed by accordingly adjusting the pressure of wheel cylinders and thereby regulating the braking force of each road wheel. To perform said acceleration slip control, the drive force of a drive wheel is controlled by regulating the braking force of said wheel if slippage occurs at said wheel. Since squat, i.e. a sudden sinking of the rear part of the vehicle body in case of a rear-wheel-drive vehicle occurs if the braking force is acting on the driven rear wheels to reduce slippage thereof, the suspension characteristic of said rear wheels is altered by a controller to a harder state. This altering of the suspension characteristic may be achieved by increasing the spring constant of a suspension, the damping force of a shock absorber or the stiffness of a suspension bush or stabilizer. This known suspension control system is of complicate construction because wheel slippage during braking as well as during acceleration is controlled by means of braking force applied to the corresponding road wheels, and additionally the suspension characteristic of the such braked road wheels is hardened to restrain the aftermath of this type of wheel slippage control. It is therefore an object of the present invention to provide a suspension control system of the above mentioned type which achieves higher stability of the vehicle by providing improvement for the aforementioned unsolved drawback in the prior art. This object is achieved by the characterizing features of claim 1. Advantageous embodiments are described in the subclaims. BRIEF DESCRIPTION OF THE DRAWINGS 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 a circuit diagram of a hydraulic active suspension system for which the preferred embodiment of a suspension control system is applicable; Fig. 2 is a section of one example of a pressure control valve to be employed in the hydraulic active suspension system of Fig. 1; Fig. 3 is a schematic block diagram of the preferred embodiment of the suspension control system according to the invention; Fig. 4 is a flowchart of a wheel load distribution controlling routine to be executed by the preferred embodiment of the suspension control system according to the invention; and Fig. 5 is a chart showing relationship between the wheel slippage and driving and braking force; Fig. 6 is a circuit diagram of a pneumatic suspension system for which the preferred embodiment of a suspension control system is also applicable; and Fig. 7 is a diagrammatic illustration of a variable stabilizer for which the preferred embodiment of a suspension control system is also applicable; DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now to the drawings, it should be initially noted that the preferred embodiment of a suspension control system, according to the present invention, is applicable for various constructions of vehicular suspension systems which have capabilities of adjusting of suspension characteristics and whereby have capability of adjusting wheel load distribution. Therefore, it should be clearly appreciated that the suspension systems discussed herebelow are mere examples of application of the preferred embodiment of the suspension control system according to the invention. Referring to Fig. 1, a hydraulic active suspension system includes hydraulic cylinders 10FL, 10FR, 10RL and 10RR for respective of front-left, front-right, rear-left and rear-right wheels. As can be appreciated, each hydraulic cylinder 10FL, 10FR, 10RL and 10RR is connected to a vehicle body (not shown) at the top of a piston rod 10c rigidly connected to a piston 10h and to a suspension member (not shown) rotatably supporting corresponding one of front-left, front-right, rear-left and rear-right wheels and pivotably supported on a vehicle body. As can be seen, the piston divides the interior space of the hydraulic cylinder in to upper and lower fluid chambers 10g and 10e. Each of the hydraulic cylinders 10FL, 10FR, 10RL and 10RR has a communication port 10a for establishing fluid communication between the lower fluid chamber 10e and a pressure accumulator 11 via a flow restriction orifice 12. On the other hand, the upper and lower fluid chambers 10g and 10e are communicated with each other via an axially extending flow restriction orifice 10f. The orifices 12 and 10f serve for restricting fluid flow therethrough for inducting fluid pressure differences between the upper and lower fluid chambers 10g and 10e and thus generate damping force. For respective of the upper chambers 10g of the hydraulic cylinders 10FL, 10FR, 10RL and 10RR, control pressures Pc₁, Pc₂, Pc₃ and Pc₄ are introduced from a hydraulic pressure control circuit which will be discussed later. The control pressure Pc₁, Pc₂, Pc₃ and Pc₄ serve for adjusting fluid pressure in the upper fluid chamber 10g. The adjusting fluid pressure in the upper fluid chamber 10g is gradually transferred to the lower fluid chamber 10e via the flow restriction orifice 10f. According to variation of the fluid pressure in the upper fluid chamber 10g, the piston 10h strokes up and down according to pressure difference between the upper and lower fluid chambers 10g and 10e to establish hydraulic force balance between the forces exerted on the upper and lower sides of the piston. Since the piston carries the piston road 10c, the relative distance between the vehicle body and the suspension member is adjusted according to the piston stroke. Therefore, by adjusting the control pressure Pc₁, Pc₂, Pc₃ and Pc₄ to be introduced into the upper fluid chamber 10g, the relative distance between the vehicle body and the suspension member can be appropriately adjusted. On the other hand, the pressure accumulator 11 as well as the orifices 12 and 10f serves for absorbing pressure variation and damping movement of the piston and thus effective for absorbing high frequency and low amplitude vibration energy such as that induced by road shock. As seen from Fig. 1, the hydraulic control circuit includes a fluid pump 13, one-way check valves 14, 15 and 16, a pressure accumulator 17, a pressure relief valve 18, an oil cooler 19, a fluid reservoir 20 and a shut-off valve 21. These components form a pressure source cirucit in the hydraulic control circuit for generating hydraulic pressure and distributing generated hydraulic pressure to respective of the hydraulic cylinders 10FL, 10FR, 10RL and 10RR. Namely, while the fluid pump 13 is driven, the hydraulic pressure is circulated through a hydraulic circuit. The maximum line pressure of a supply line is determined by a set pressure in the pressure relief valve 18 by draining excessive pressure to the fluid reservoir 20 via a drain line. On the other hand, the shut-off valve 21 is associated with a main power switch, such as an ignition switch, so that the valve is held in open position while the main power switch is ON and responsive to turning OFF to switch the valve into shut-off position for blocking fluid communication between the pressure source circuit and the hydraulic cylinders 10FL, 10FR, 10RL and 10RR for maintaining the fluid pressure therein. Proportioning solenoid valves 23FL, 23FR, 23RL and 23RR are provided between the pressure source circuit and the hydraulic cylinders 10FL, 10FR, 10RL and 10RR for generating control pressures Pc₁, Pc₂, Pc₃ and Pc₄ to be supplied to the hydraulic cylinders, which valves therefore serve as pressure control valve. Each of the proportioning solenoid valves 23FL, 23FR, 23RL and 23RR incorporates a proportioning solenoid 23a for adjusting pilot pressure in the pressure control valves 23FL, 23FR, 23RL and 23RR. Fig. 2 shows one example of the pressure control valve applicable for the shown construction of the active suspension system. In the construction of the pressure control valve shown in Fig. 2, the pressure control valve 23 includes a valve housing 142. The valve housing 142 defines a valve bore 144 which is separated in to a valve chamber 142L and a control chamber 142U by means of a partitioning wall 146. The partitioning wall 146 is formed with a communication path opening 146A for communication between the control chamber 142U and the valve chamber 142L. As seen from Fig. 2, the control chamber 142U and the valve chamber 142L are arranged in alignment to each other across the communication path opening 146A. In parallel to a section of the partitioning wall 146 extending perpendicular to the axis of the valve chamber 142L and the control chamber 142U, a fixed orifice defining partitioning member 150 is provided. The partitioning member 150 defines a throttling orifice which is oriented substantially in alignment with the communication path opening 146A. The partitioning wall 146 and the partitioning member 150 are cooperative for defining a pilot chamber PR therebetween. A valve spool 152 is thrustingly and slidingly disposed within the valve chamber 142L. The valve spool 152 defines an upper feedback chamber FU between the tip end thereof and the partitioning member 150. The valve spool 152 also defines a lower feedback chamber FL between the other tip end thereof and the bottom of the valve chamber 142L. Offset springs 160U and 160L are disposed within the upper and lower feedback chambers FU and FL, which offset springs exerts spring force to the valve spool 152 for resiliently restricting movement of the latter. Resilient force of the offset springs 160U and 160L are so set as to balance to place the valve spool 152 at a neutral position, when the fluid pressure in the upper and flow feedback chambers FU and FL balances to each other. The valve chamber 142L is communicated with a supply line 35 via an supply port 154s, the drain line via a drain port 154r and a pressure control line 38 via a control port 154o, which supply port, drain port and control port are defined in the valve housing 142. The valve spool 152 at the aforementioned neutral position, blocks fluid communication of the pressure control chamber PC with any of the supply port 154s and the drain port 154r. As a result, as long as the valve spool 152 is maintained at the neutral position, overall fluid force in the hydraulic circuit downstream of the pressure control valve, which circuit includes the upper fluid chamber 10g of the hydraulic cylinder 10 is held constant. The valve spool 152 is formed with lands 152a and 152b connected to each other via smaller diameter bar-like section 152e. The land 152a is oriented adjacent the lower feedback chamber FL so as to subject the tip end to the fluid pressure in the lower feedback chamber. Similarly, the land 152b is oriented adjacent the upper feedback chamber FU so as to subject the tip end to the fluid pressure in the upper feedback chamber. The bar-like section 152e between the lands 152a and 152b is cooperative with the peripheral wall of the valve chamber 142L in order to define therebetween a pressure control chamber PC. A fluid flow path 152d is formed through the valve spool 152. The fluid flow path 152d has one end communicated with the pressure control chamber PC and the other end communicated with the lower feedback chamber FL. A fixed flow restricting orifice 152f is formed in the fluid flow path 152d for restricting fluid flow therethrough. A poppet valve member 148 is disposed within the control chamber 142U for thrusting movement therein. The poppet valve member 148 has a valve head 148a of an essentially conical configuration. The valve head 148a opposes to the communication path opening 146A of the partitioning wall 146. The poppet valve member 148 is operably associated with a proportioning solenoid assembly 29 as the actuator. The proportioning solenoid assembly 29 comprises a housing 162 rigidly secured on the valve housing 142 and defining an internal space to receive therein a plunger 166. The plunger 166 has a plunger rod 166A. The tip end of the plunger rod 166A is kept in contact with the tip end of the poppet valve member 148 remote from the valve head 148a. Therefore, the poppet valve member 148 is axially driven by means of the plunger 166 to control the path area in the communication path opening 146A according to the position of the tip end of the plunger rod 166A. Adjusting of the path area in the communication path opening 146A results in variation of fluid pressure to be introduced into the pilot chamber PR. In order to control the position of the plunger 166 with the plunger rod 166A, a proportioning solenoid coil 168 is housed within the housing 162 and surrounds the plunger 166. The interior space of the housing 162 is connected to the control chamber 142U for fluid communication therebetween. The plunger 166 is formed with a fluid path 166B for fluid communication between upper and lower sections of the interior space. Therefore, the fluid pressure in the upper and lower sections of the interior space of the housing 162 is held equal to the pressure in the control chamber 142U. This cancels fluid pressure to be exerted on the poppet valve and the plunger so that the position of the tip end of the plunger rod 166A can be determined solely depending upon the magnitude of energization of the proportioning solenoid coil 168. As seen from Fig. 2, the poppet valve member 148 has a cylindrical larger diameter section 148b for separating the control chamber 142U into upper section and lower section 142Uu and 142Ul. The upper and lower sections 142Uu and 142Ul are communicated with the drain port 154r via a pilot return path PT. A multi-stage orifice Pr is provided in the pilot return path PT for restricting fluid flow therethrough. The multi-stage orifice Pr comprises a plurality of strips formed with through openings and is so designed that one of the orifice oriented at most upstream side is mainly effective for restricting fluid flow when fluid flowing therethrough is steady flow and that all of the orifices of respective strips are equally effective for restricting fluid flow when fluid flow therethrough is disturbed and not steady. Therefore, as will be appreciated herefrom, the multi-stage orifice Pr employed in the shown embodiment serves to provide greater fluid flow restriction against non-steady or disturbed fluid flow than that for the steady flow. As seen from Fig. 2, the multi-stage orifice Pr is provided upstream of the upper and lower sections 142Uu and 142Ul. On the other hand, a fixed throttling orifice Pd is provided at an orientation downstream of the lower section 142Ul and upstream of the upper section 142Uu. Similarly, the pilot chamber PR is communicated with the supply port 154s via a pilot path PP. A multi-stage orifice Qp which has similar construction and flow restricting function to that of the multi-stage orifice Pr is provided in the pilot path PP. A fixed throttle orifice Pro is also provided in the drain port 154r for restricting fluid flow therethrough. The diameter of the fluid path at the orifice Pro is so selected as to create great flow restriction against pulsatile fluid flow cyclically varying the fluid pressure at a frequency approximately 1 Hz. As can be seen from Fig. 2, the pressure control valve 23 is so arranged as to direct the axis of the valve bore 144 parallel to the longitudinal axis of the vehicle body. The longitudinal acceleration to be exerted on the vehicular body is much smaller than the lateral acceleration and vertical acceleration exerted on the vehicle body. Therefore, by arranging the pressure control valve 23 so that the poppet valve 148 and the valve spool 152 thrustingly move in longitudinal direction, influence of the externally applied acceleration can be minimized. As seen from Figs. 1 and 3, the pressure control valves 23FL, 23FR, 23RL and 23RR, a control unit 26 is provided in order to determine control signals I₁, I₂, I₃ and I₄ for the actuators of respective pressure control valves according to a preselected suspension control parameters. In the shown embodiment, a vehicle speed V and a steering angle displacement ϑ are selected as suspension control parameters. For providing the vehicle speed data V and the steering angular displacement data ϑ, a vehicle speed sensor 27 and a steering angle sensor 28 are connected to the control unit 26. As will be appreciated, based on the steering angle data ϑ and the vehicle speed data V a data representative of vehicular attitude change can be derived. The control unit 26 comprises an input inter face 33, CPU 34, ROM 35, RAM 36 and an output interface 37. For the input interface, wheel speed sensors 29, 30, 31 and 32 respectively designed for monitoring rotation speed of respectively corresponding one of front-left, front-right, rear-left and read-right wheels for producing front-left, front-right, rear-left and rear-right wheel speed indicative signals Vw₁, Vw₂, Vw₃ and Vw₄. The control unit 26 derives a vehicle speed representative values V₁, V₂, V₃ and V₄ on the basis of the vehicle speed data V and the steering angle data ϑ. Based on the vehicle speed representative values V₁, V₂, V₃ and V₄ and the wheel speed indicative signals Vw₁, Vw₂, Vw₃ and Vw₄, the control unit 26 derives wheel slippage S₁, S₂, S₃ and S₄ utilizing the following equations: During Acceleration Si = 1 - Vi/(r₀ x Vw i ) (1) During Deceleration (Braking) Si = 1 - (r₀ x Vw i )/Vi (2) wherein r₀ is tire load diameter; i is 1, 2, 3 and 4 The control unit 26 performs tire/road traction dependent suspension control with taking the wheel slippage S₁, S₂, S₃ and S₄ as the tire/road friction representative data. The preferred process of tire/road traction dependent suspension control will be discussed herebelow with reference to the flowchart in Fig. 4. As will be appreciated, the control unit 26 in the automotive vehicle performs various control operations, such as engine control, brake control, suspension control and so forth. Therefore, the shown routine in Fig. 4 is triggered at every predetermined timing, such as every 20 ms. At a step 101, the vehicle speed indicative data V, the steering angle data ϑ are read out. Based on the vehicle speed indicative data V and the steering angle data ϑ and with taking specification of the vehicle for which the suspension control system is applied, the vehicle speed data V₁, V₂, V₃ and V₄ at respective of the front-left, front-right, rear-left and rear-right wheels are derived as references valves for deriving wheel slippage at respective wheels, at a step 102. Subsequently, the wheel speed indicative data Vw₁, Vw₂, Vw₃ and Vw₄ are read out at a step 103. Based on the vehicle speed data V₁, V₂, V₃ and V₄ and the wheel speed indicative data Vw₁, Vw₂, Vw₃ and Vw₄, wheel slippages S₁, S₂, S₃ and S₄ are derived according to one of the equations (1) and (2), at a step 104. Based on the wheel slippage data S₁, S₂, S₃ and S₄ derived at the step 104, the wheel having the maximum wheel slippage S max is selected at a step 105. Then, the maximum wheel slippage indicative value S max is compared with a predetermined wheel slippage threshold S limit at a step 106. When the maximum wheel slippage S max is smaller than or equal to the wheel slippage threshold S limit , judgement is made that wheel slippage is not substantial and within the acceptable range. If this is the case, process directly goes END. Therefore, normal suspension control irrespective of the wheel slippage is performed in another routine. The process of suspension control in another routine has been known in various ways. On the other hand, when the maximum wheel slippage S max as compared at the step 106 is greater than the wheel slippage threshold S limit , process goes to a step 107 in order to perform Wheel slippage dependent suspension control. In the wheel slippage dependent suspension control, wheel load distribution is adjusted so as to reduce the wheel slippage in order to maintain the maximum wheel slippage smaller than the wheel slippage threshold, For this purpose, in the wheel speed dependent suspension control, the wheel load for the wheel having the maximum wheel slippage is increased in a magnitude derived for reducing the maximum wheel slippage down to the wheel slippage threshold. It is general process of control in the active suspension system to independently control the wheel load for respective of individual wheels and to maintain wheel load difference in longitudinal direction and lateral direction equal to each other for maintaining the yawing moment unchanged. On the other hand, the relationship between the wheel slippage S i and the driving and braking forces is shown in Fig. 5. In Fig. 5, the curve A shows variation of wheel slippage S i versus the driving and braking force Q i when the wheel load distribution is not adjusted for exerting wheel load W. On the other hand, the curve B shows variation of the wheel slippage S i versus the driving and braking force Q₁ when the wheel load distribution is adjusted so that the wheel load on the wheel having the maximum wheel slippage is increased for exerted wheel load W′ (W + ΔW) by the process of the preferred process. Assuming the wheel load is W and the driving and braking force is Qi₁, the wheel slippage at the point X becomes Si₁ as shown in Fig. 5. If this wheel slippage Si₁ is greater than the wheel slippage threshold S limit , cornering force CF is rapidly decreased to cause breakage of tire/road traction. Breaking of traction is at first caused at the wheel having the maximum wheel slippage. This breaks balance of the vehicular body to lead attitude change in the vehicle body. At this time, by reducing the wheel slippage from Si₁ to Si₂ which is smaller than or equal to the wheel slippage threshold as pointed by the point Z, the driving and braking force is undesirably reduced to Qi₂. Therefore, according to the present invention, by adding ΔW to the normal wheel load W, wheel slippage is reduced from Si₁ to Si₂ without causing drop of the driving and braking force Qi₁ as pointed at the point Y. As can be appreciated, since the wheel slippage dependent suspension control is performed not only when the wheel slippage is caused by longitudinal or lateral acceleration, but also when the friction coefficient µ at the road surface is different in each individual wheel, vehicular stability on the low friction road can be assured. Fig. 6 shows a diagram of a pneumatic suspension system for which the preferred embodiment of the suspension control system is also applicable. In this embodiment, front-left, front-right, rear-left and rear-right suspension systems 58FL, 58FR, 58RL and 58RR have variable volume pneumatic chamber 58a. Each of the pneumatic chambers 58a is connected to a pressurized air source including an electric motor 50, a motor driven compressor 51 incorporating an air induction associated therewith an air filter 52, an air dryer 53. For the pressure supply line connected to the discharge outlet of the compressor via the air dryer 53, a pressure reservoir tank 55 is connected via an one-way check valve 54. A main valve 57 is also connected between the pressure supply line and the pressure reservoir tank 55. Pressure control valves 56FL, 56FR, 56RL and 56RR are provided in respective of branch line connecting the pressure supply line to the pneumatic chambers 58a of respective suspension system 58FL, 58FR, 58RL and 58RR. For respective of the pneumatic chambers 58a, sub-tanks 61 is connected via cut valves 60. Each of the pressure control valves 56FL, 56FR, 56RL and 56RR comprises a three way valve variable of valve position between supply position for supplying air pressure Pc₁, Pc₂, Pc₃ and Pc₄ to corresponding one of the branch lines, discharge position for ventilating or discharging the air pressure in the pneumatic chamber 58a and shut-off position for blocking air communication through the branch line. As can be seen from Fig. 6, each pressure control valve 56FL, 56FR, 56RL and 56RR has a pair of electromagnetically operable solenoids 56a and 56b. In order to control operation of the motor 50, the main valve 57, the pressure control valves 56FL, 56FR, 56RL and 56RR, and the cut valves 60, a microprocessor based control unit 26 is performed, which control unit implements the preferred suspension control according to the invention, substantially in the same manner as that in the former embodiment. Similarly to the former embodiment, the control unit 26 is connected to the vehicle speed sensor 27, the steering angle sensor 28, and the wheel speed sensors 29, 30, 31 and 32. In addition, the control unit 26 is also connected to pressure sensors 59 which monitors air pressure in respective branch lines. The control unit 26 is further connected to stroke sensors 62 which monitors relative distance between the vehicle body and the suspension member in respective suspension systems 58FL, 58FR, 58RL and 58RR. In the shown construction, the wheel load distribution on each of the front-left, front-right, rear-left and rear-right wheels is controlled by adjusting air pressure in respective of the pneumatic chambers 58. The pressure in the pneumatic chamber 58a is primarily controlled by the pressure control valves 56FL, 56FR, 56RL and 56RR by switching valve positions. The sub-tank 61 and the cut valves 60 are cooperative with the pressure control valve for adjustment of the air pressure in the pneumatic chambers. In general, such pneumatic suspension system controls air pressure in the pneumatic chambers 58a in order to regulate the vehicular body height level on the basis of the vehicular height indicative data supplied from the stroke sensors 62 so as to maintain the vehicular height within a predetermined height range. Therefore, the control unit 26 normally perform height adjustment operation depending upon the vehicular body height indicative signals of the stroke sensors 62. On the other hand, the control unit controls operation of the motor 50, the main valve 57 and cut valves 60 on the basis of the pressure indicative signal from the pressure sensors 59. The control unit 26 further performs the similar process as that discussed with respect to Fig. 5. Namely, when the maximum wheel slippage S max is greater than the wheel slippage threshold S limit , wheel slippage dependent suspension control is performed so as to increase the air pressure in the pneumatic chamber of one of the suspension system 58FL, 58FR, 58RL and 58RR is a magnitude ΔW sufficient for reducing wheel slippage to be smaller than or equal to the wheel slippage threshold. Though the specific construction of the pneumatic suspension has been discussed hereabove, the shown embodiment of the suspension control system is applicable for various constructions of the pneumatic suspension systems. Fig. 7 diagrammatically shows a variable stabilizer, for which the preferred embodiment of the suspension control system according to the present invention is applicable. In the shown construction, the variable stabilized comprises a pair of hydraulic cylinders 70 and 71 generally provided in front and rear suspension systems. The hydraulic cylinder 70 defines an upper fluid chamber 70a and a lower fluid chamber 70b. On the other hand, the hydraulic cylinder 71 defines an upper fluid chamber 71a and and a lower fluid chamber 71b. The upper fluid chamber 70a of the hydraulic cylinder 70 is connected to the lower fluid chamber 71b of the hydraulic cylinder 71 via a communication line 72. A pressure accumulator 73 is connected to the communication line 72. On the other hand, the upper fluid chamber 71a of the hydraulic cylinder 71 is connected to the lower fluid chamber 70b of the hydraulic cylinder 70 via a communication line 74. A pressure accumulator 75 is connected to the communication line 74. With the shown construction, assuming that the upward force ΔF f is exerted on the front wheel and a piston 70c of the hydraulic cylinder 70 strokes in a magnitude ΔX f , upward force Δf (approximately corresponding to the upward force ΔF f ) is exerted on a piston 71c of the hydraulic cylinder 71 if the piston stroke of the piston 71c is zero. By this stabilization effect can be obtained. In the shown construction, when the vehicular driving condition is close to the critical point due to braking or cornering, and assuming wheel slippage is caused due to difference of road friction or unbalance of the wheel load on the wheels, downward force to be exerted on the hydraulic cylinders 70 and 71 is reduced. In such case, stabilization effect of the variable stabilizer similar to that in response to the upward force ΔF f can be obtained. In the shown embodiment, wheel slippage dependent control can be performed for adjusting the wheel load for the wheel having the maximum wheel slippage by adjusting set pressure P f and ΔP r of the pressure accumulators 73 and 75. By this, the equivalent effect for assurance of vehicle driving stability can be obtained. Therefore, the present invention fulfills all of the objects and advantages sought therefor. 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 example, though the shown embodiments have been directed to adjustment of the hydraulic or pneumatic pressure to be for active cylinders, it may be possible to adjust damping characteristics of variable shock absorber. Therefore, the suspension control system according to the present invention may be applicable for suspension systems employing variable damping characteristics shock absorber.

Claims: A suspension control system for a vehicle having a vehicle body and a plurality of road wheels, comprising: a suspension system (10; 58; 70,71) disposed between the vehicle body and each of said road wheels; first means (23; 56; 73,75) associated with said suspension system (10; 58; 70,71) for adjusting the load (W) exerted on each of said road wheels; second means (27,28,62) for monitoring preselected suspension control parameters for providing a first set of parameter data (V,ϑ); third means (27-32) for monitoring the state of wheel slippage at a plurality of said road wheels for providing a second set of parameter data (V,ϑ,Vw i ); and fourth means (26) receiving said first and second set of parameter data (V,ϑ; V,ϑ,Vw i ), for controlling operation of said first means (23; 56; 73,75) and for deriving a wheel slippage indicative value (S i ) for each of said plurality of road wheels based on said second set of parameter data (V,ϑ,Vw i ), characterized in that, if the maximum wheel slippage indicative value (S max ) of said plurality of wheel slippage indicative values (S i ) is less than or equal to a predetermined wheel slippage threshold limit (S limit ), normal suspension control based on said first set of parameter data (V,ϑ) is carried out by said fourth means (26), and if said maximum wheel slippage indicative value (S max ) is greater than said threshold limit (S limit ), said fourth means (26) controls said first means (23; 56; 73,75) such that the wheel load (W) exerted on the wheel having the maximum wheel slippage is increased by a magnitude (ΔW) sufficient for reducing wheel slippage to be smaller than or equal to said wheel slippage threshold limit (S limit ). A suspension control system according to claim 1, characterized in that said suspension system comprises a variable fluid chamber (10g,58a) for adjusting suspension characteristics; a pressure control circuit including a fluid pressure source (13,51) and said first means (23,56) is connected to said variable fluid chamber (10g, 58a); and said first means comprises a pressure control valve (23,56) which controls supplying and draining the fluid pressure (Pc i ) in said variable pressure chamber (10g,58a). A suspension control system according to claim 2, characterized in that, if said maximum wheel slippage indicative value (S max ) is less than or equal to said threshold limit (S limit ), said fourth means (26) supplies a control signal (I i ) for that pressure control valve (23,56) which corresponds to said wheel having the maximum wheel slippage, in order to increase the fluid pressure (Pc i ) in the corresponding variable fluid chamber (10g,58a) by an additional magnitude derived on the basis of said first set of parameter data (V,ϑ). A suspension control system according to claim 3, characterized in that said fourth means (26) derives said additional magnitude of pressure depending upon the level of wheel slippage in relation to said threshold limit (S limit ). A suspension control system according to one of claims 2 to 4, characterized in that said suspension system is a hydraulic active suspension system including a hydraulic cylinder (10) having a variable pressure chamber (10g) as said variable fluid chamber, that said fluid pressure source comprises a hydraulic pressure source (13) and that said pressure control valve comprises a proportioning solenoid valve (23). A suspension control system according to one of claims 2 to 4, characterized in that said suspension system is a pneumatic suspension system (58) including a variable volumen pneumatic chamber (58a) as said variable fluid chamber, that said fluid pressure source comprises a pressurized air source (51) and that said pressure control valve is a three way-three position valve (56). A suspension control system according to claim 1, characterized in that said suspension system comprises a variable stabilizer comprising a pair of first and second cylinders (70,71), said first cylinder (70) defining an upper first and a lower second chamber (70a,70b) opposing across a first thrusting piston (70c), said second cylinder (71) defining an upper third and a lower fourth chamber (71a,71b) opposing across a second thrusting piston (71c), said first chamber (70a) being communicated with said fourth chamber (71b) via a first line (72), said second chamber (70b) being communicated with said third chamber (71a) via a second line (74) and said first means (73,75) being provided in said first and second lines (72,74) for adjusting set pressure (ΔP f ,ΔP r ) therein. A suspension control system according to claim 7, characterized in that said first means comprises a pressure accumulator (73,75) variable of set pressure (ΔP f ,ΔP r ) thereof, said pressure accumulator (73,75) being connected to each of said first and second lines (72,74).

Inventors: IMASEKI TAKASHI, SUGASAWA FUKASHI, YAMAMURA TOMOHIRO, YOKOTE MASATSUGU, IMASEKI, TAKASHI, SUGASAWA, FUKASHI, YAMAMURA, TOMOHIRO, YOKOTE, MASATSUGU
Applicants: NISSAN MOTOR, NISSAN MOTOR CO., LTD.
Application year: 1990
Main classification: B60G17
IPC codes: B60G21, B60G17