VEHICLE SUSPENSION ARRANGEMENTS

Abstract: A suspension system for a vehicle road wheel comprises a wheel suspension device (1) having a double-acting hydraulic actuator and a gas spring in parallel, the gas spring being controllable to provide a selectively variable offset, sensor means (6, 90) arranged to sense load changes experienced by the suspension device and to provide an output dependent thereon, and control means (95) respon­ sive to the sensor means output to effect adjustment of the hydraulic actuator pending change of the offset of the gas spring means to accommodate the load changes. Also disclosed is a damping system for damping movement of a piston (64) within a cylinder, the damping system comprising a valve (80) operable to connect spaces (70, 74) within the cylinder at opposite sides of the piston to fluid pressure and/or return lines (84, 85) so as to apply selected positive or negative damping to the movement of the piston. The damping system can be incorporated in the suspension system so that the piston (64) of the hydraulic actuator is damped, under control of signals from the control means (95) applied to the servo-valve (80).

Description: The invention relates to vehicle suspension arrangements, in particular to active vehicle suspension arrangements in which the wheel position relative to the vehicle is adjusted in response to sensed conditions. In European Patent Application 84 300 370.8 (Publication EP 0114 757 A), there is described a vehicle suspension system comprising wheel suspension devices the displacement of which can be varied, and means responsive to the loading of the suspension devices to feed back to the devices electric signals effecting a predetemined displacement. In European Patent Application 84 307 288.5 (Publication EP 0 142 947A), there is disclosed a wheel suspension device for use in such a system and comprising an hydraulic actuator of which the piston position is adjustable in response to forces acting thereon and/or to a selected control input, and a load bearing strut responsive to the actuator piston experience to minimise the load carried by the actuator. The suspension system of EP 0 114 757 A offers substantial advantages over previous passive suspension systems, for example self-levelling of the vehicle under varying loads, due either to payload changes or to externally applied loads, can be effected. The system moreover has the ability to stablilise the vehicle under inertial loads such as braking, accelerating and cornering, so as to prevent changes in attitude of the vehicle. Additionally, the damping characteristics can be adaptively varied according to dynamic requirements of both the sprung and unsprung masses, to minimize transmitted accelerations, while maximising the grip of the wheel on the road under varying conditions. Also, the system can be adapted to receive a selected energy input to further modify the damping characteristics. German patent specification no. 2844413 also discloses a suspension system having wheel suspension devices each consisting of an actuator and a gas spring. The displacement of the actuator can be varied automatically by a control system to stablize the vehicle under lateral acceleration and the gas spring can be operated manually to adjust the level of the vehicle. Furthermore, JP-A-59 186 716 discloses suspension systems having shock absorbers including damping force variable dampers adjustable through first solenoid valves, and spring forces variable rate air springs adjustable through second solenoid valves, the second solenoid valves being operable with a predetermined time lag after operation of the first solenoid valves. Such operation with a time lag is provided only when a difference between an existing steering angle and a weighted mean value derived from a steering angle detector exceeds a predetermined value, and its purpose is to improve driving comfort. The full advantages of such an active vehicle suspension system are obtained only by the input of considerable energy, and it would accordingly be advantageous to provide an active vehicle suspension system which reduces this energy demand under a wide range of conditions, for a given set of characteristics. In a system embodying the invention, as defined in claim 1 the actuating means can comprise a double-acting hydraulic actuator connected to a source of pressure fluid through a servo-valve responsive to signals generated in response to the load experience of the device, in parallel with a 2-port electrically controlled proportional valve. The gas spring can comprise a selectively variable volume of fluid in the device either in contact with a body of gas also within the device or associated with an external hydraulic accumulator. The volume of the fluid body is adjustable by means of a 3-port valve which arranged to respond to the pressure difference across the actuator piston, sensed either electrically or hydraulically or to the load experienced by the device. Under the steady loads, that is, when the mean pressure across the actuator piston is zero, the damping of the device is primarily controlled by the valve in parallel with the servo-valve, for example, the 2-port valve mentioned, either passively, or if the valve is electrically controlled, adaptively, the valve being controlled to effect any change in damping characteristics called for externally of the device. Sudden changes in the steady load carried by the device, for example, a payload change or an inertial load change, are controlled by adjustment of the actuator, as by way of the solenoid valve, until the gas spring offset had been changed, as through the 3-port valve. However embodied, the invention provides that the energy consumption of the system is a function only of the active component of the dynamic movement of the device and the flow through the actuator valving arrangement caused by the pressure difference across the actuator piston, whilst the gas spring offset is being adjusted to bring this pressure difference to zero. Substantial reduction in energy demand is achieved over the demand in conventional systems which the demand is a function of the total motion of the suspension devices. The invention can also provide for selective damping of the movement of the piston in the actuator cylinder. This damping can be provided by valve means selectively movable to communicate the cylinder interiors on opposite sides of the piston with pressure fluid supply and return lines. The movement of the piston can thus be damped by fluid pressure either positively or negatively, and to a selected degree. The invention is further described below, by way of illustration, with reference to the accompanying drawings, in which: Figures 1A and 1B together are a sectional front view of a vehicle wheel suspension device for use in vehicle suspension systems embodying the invention, shown in an extended position; Figure 2 schematically shows the suspension device of Figure 1 incorporated in a first vehicle wheel suspension system in accordance with the invention; Figure 3 schematically shows the suspension device of Figure 1 incorporated in a second vehicle wheel suspension; Figures 4, 5 and 6 schematically show how a servo-valve included in the system can be operated to achieve desired characteristics; and Figure 7 graphically displays the relationship of the velocity of the piston of an actuator included in the system to the load across the piston. The vehicle wheel suspension device 1 illustrated in Figures 1 and 2 comprises a lower end fitting 2 formed with a loop 4 for connection of the device to a vehicle axle assembly (not shown). At its upper side, the fitting 2 has a recess accommodating a load cell 6 operative between the fitting and a head member 8. The head member 8 has a stepped recess formed in its upper side within which are received three concentric tubes 14, 22, 28, each sealed to the head member by an external O-ring or like seal. The tube 22 is spaced from the tubes 14 and 28 to define respective intermediate spaces 20 and 26. The assembly comprising the tubes 14, 22, 28 is secured to the head member 8 by means of a steel split collar 32. At its upper end, the inner tube 14 is sealingly secured to an extension portion 34 of the outer tube 28, the upper end of which is closed. A ring 38 makes a sliding seal to the exterior of the extension portion 34 by means of sealing rings 39. An upper end fitting 42 has a lower tubular end portion 40 secured to the ring 38. The fitting 42 is closed by a reduced diameter portion 44 at its upper end, the portion being tapped to receive a threaded end element 45 for connection of the device 1 to a vehicle chassis. An outer ring 54 is secured around the lower end of the end portion 40 by means of a screw threaded connection. Secured within the lower end of the ring 54 is a sleeve 56, the upper end of which abuts the end portion 40 and the ring 38. The sleeve 56 is thus a downward extension of the end fitting 42 and at its lower end it makes a sliding seal on the exterior of the outer tube 28 by means of sealing elements 58 held by an end ring 60 secured to the sleeve end by an outer ring 62. The outer tube 28 has, below the extension tube 34, a larger diameter portion 64, which makes a sliding seal, with the sleeve 56 by means of sealing elements 66, and which functions as a piston. The space 20 between the inner tube 14 and the intermediate tube 22 communicates through a passage 24 in the head member 8 with valve means described with reference to Figures 2-6, and the outer annular space 26 between the intermediate tube and the stainless steel outer tube 28 communicates with the valve means through a passage 30 in the head member. Ports 68 at the lower end of the piston portion 64 communicate the space 26, between the tubes 22 and 28, with the space 70 between the tube 28 and the sleeve 56. Also, ports 72 at the upper end of the portion 64 communicates the space 20 between the tubes 14 and 22 with the space 74 between the extension tube 34 and the sleeve 56. The device 1 can thus function as a double acting hydraulic actuator. Pressure fluid introduced into the space 70 from a suitable pressure fluid source, by way of the valve means, passage 30, annular space 26 and ports 68, will act on the end faces of the space 70 to shorten the device. Pressure fluid supplied to the space 74 through the valve means, passage 24, annular space 20 and ports 72 will act within the space 74 to lengthen the device. Also, the interior of the end fitting 42 forms a working chamber 46 to which fluid can be supplied under pressure through a port 47. The volume of the body of hydraulic fluid in the working chamber 46 can thus be selectively adjusted to alter the load deflection characteristics of the device 1. The condition of the device 1 is sensed not only by the load cell 6 but also by a linear variable displacement transformer 90, of which the stator portion is received in a housing secured to the fitting 2 and to the head member 8, whilst the movable core or yoke portion is secured to the sleeve 56. The output of the transformer 90 is thus dependent on the length, and variations in the length, of the device 1, as represented by the position and movement of the piston portion 64 in the sleeve 56. The transformer 90 thus supplies signals indicative of at least one of the velocity, acceleration and position of the piston portion 64. In Figure 2, the port 47 of the working chamber 46 is connected through a 3-port valve 49 with a fluid pressure source 50 by way of an hydraulic accumulator 51. Figure 2 also schematically shows valve means communicating with the passages 24 and 30 in the head member 8. The valve means comprise a servo-valve 80 communicating with a fluid pressure source 86 which can, but need not be, a source separate from the source 50, by way of pressure fluid supply and return lines 84 and 85. The servo-valve 80 is connected in parallel with a 2-port electrically controlled proportional valve 82, the proportional valve in turn communicating with the head member passages 24 and 30. The outputs of the load cell 6 and of the linear variable displacement transformer 90 are supplied to control circuitry 95 arranged to generate control signals which are supplied to the valves 49, 80 and 82. The control circuitry 95 is generally of the kind described in the above referenced Publication EP 0 114 757 A. However, the control circuitry and also the characteristics of the valves of the present system are so selected and arranged that during a steady load condition, in which the mean pressure differences across the piston portion 64 is zero, the damping of the device 1 is controlled primarily by the 2-port valve 82. This control may be passive or adaptive, as by being proportional to the electrical control signal supplied by the control circuitry 95. The circuitry 95 is arranged to supply to the 3-port valve 49 control signals dependent on the load experienced by the device 1, or the pressure difference across the piston portion 64. Sudden changes in the load experienced by the device 1, which may be due to a change in the payload of the vehicle in which the system is incorporated and/or to inertia loads on accelerating or cornering, are accommodated by the servo-valve 80 until such time as the valve 49 responds to the control signals to adjust the volume of pressure fluid in the chamber 46 to alter the offset of the gas spring constituted by this chamber and the hydraulic accumulator 51. In the second system shown in Figure 3, parts similar to those of the system shown in Figure 2 are indicated by the same reference numerals. The second system differs from the first in respect of two modifications, either of which could be made to the first system independently. Firstly, the electrically controlled 3-port valve 49 is replaced by a 3-port valve 149 controlled hydraulically in response to the pressure difference across the piston portion 64, conveyed to the valve through pipe connections 150. Secondly, the 2-port electrically controlled valve 82 is replaced by an asymmetrical flow restrictive valve 182 fitted with blow off valves. The blow-off valves could instead be incorporated into the piston portion 64 itself. The characteristics of this valving arrangement are similar to those of damping valves in a shock absorber. The system of Figure 3 is arranged to respond to steady and changing load conditions in substantially the same way as that of Figure 2 adjustment of the position of the piston portion 64 being preferred to, or given priority over, adjustment of the volume of pressure fluid in the chamber 46. If it is desired to obtain damping characteristics requiring an energy input to the system, the control circuitry 95 is arranged to supply to the servo-valve 80 a signal to effect acceleration or deceleration, as required, of the actuator operation. Thus, Figures 4, 5 and 6 schematically show the servo-valve 80 as comprising a spool member 100 having four valve spools co-operable with ports communicating with the pressure and return lines 84 and 85 and with the fluid lines 88 and 89 extending to the valve 82, to the spaces 70 and 74 of the actuator portion of the device 1. The lines 88 and 89 communicate with the return line 85 through respective non-return valves and thus to the spaces 70 and 74 of the actuator portion of the device 1. The lines 88 and 89 communicate with the return line 85 through respective non-return valves 101, 102. In Figure 4, the position of the servo-valve spool member 100 is such as to communicate both of lines 88 and 89 with the return line 85. Communication between the pressure line 84 and lines 88 and 89 is blocked. The piston portion 64 is moving downwardly as indicated by arrow 104, so fluid flows from the return line 85 into the space 74 and to it from the space 70. Because fluid movement is to and from the return line only, the piston portion 64 is substantially decelerated or highly damped. In the position of the spool member 100 shown in Figure 5, the piston portion 64, which is moving upwardly, is accelerated, that is, it is negatively damped. This occurs because the position of the valve spools is such as to communicate the pressure line 84 with the line 89, so that the fluid pressure acts in the space 70 in the direction of movement of the piston portion. The space 74 communicates through the line 88 with the return line 85 for escape of the fluid displaced by the piston portion movement. In Figure 6, the servo-valve spool member 100 occupies a position in which neither the pressure line 84 nor the return line 85 communicate with the line 89. The line 88 communicates with the return line 85. The piston portion 64 is moving upwardly as indicated by arrow 106 and fluid is readily expelled from the space 74 through the line 88. The corresponding volume of fluid is drawn into the space 70 but only through the non-return valve 102 which communicates the return line 85 with the line 89. The piston portion 64 is therefore damped, but to a lower degree than when fluid is drawn directly from the return line 85 as in the situation shown in Figure 4. Thus it will be seen that by appropriate positioning of the spool member 100 under control of the control circuitry 95, damping and negative damping in desired degrees can be applied to the movement of the piston portion 64 in either direction. In Figure 7, the characteristics of the combination of the hydraulic actuator portion of the device 1 and the servo-valve 80 are graphically represented. The velocity of the piston portion 64 is plotted along the Y-axis against the load across the piston on the X-axis for various percentages of the rated current, the "consumed flow" being indicated by the shaded area. Although primarily intended for use in road vehicles, it will be evident that the applications of suspended systems embodying the invention are not so limited. Both the vehicle suspension system and the damping system of the invention can be embodied in various ways other than in accordance with the above description and accompanying drawings.

Claims: A vehicle suspension system comprising a wheel suspension device (1) having an adjustable double acting hydraulic actuator (56,64) in parallel with a variable offset gas spring (46,51), said actuator (56,64) having first and second spaces (70,74), valve means (80) to connect the first and second spaces with supply and return lines (84,85) for hydraulic fluid to effect adjustment of the actuator, a controller (95) to adjust the suspension device (1) in response to the output of sensor means (6,90) sensing load changes experienced by the suspension device (1), said controller (95) effects adjustment of the actuator (56,64) before adjustment of the gas spring (46, 51), and there are further valve means (82) in parallel with the valve means (80) to connect the spaces (70, 74) to effect adjustment of the actuator (56,64), said further valve means (82) being responsive to signals from the controller (95). A suspension system as claimed in Claim 1 wherein the valve means (80) comprises a servo-valve responsive to signals from the controller (95) and the further valve means (82) comprises a two-port proportional valve responsive to signals from the controller (95). A suspension system as claimed in any preceding claim wherein the gas comprises a liquid body in contact with a body of gas (51), the volume of the liquid body being adjustable to thereby adjust the variable offset of the gas spring. A suspension system as claimed in Claim 3 wherein a pressure line communicates a fluid pressure source (50) with the liquid body, and includes valve means (49) responsive to signals from the controller (95). A suspension system as claimed in Claim 4 wherein the valve means (49) comprises a three-port valve responsive to electric signals from the controller (95). A suspension system as claimed in any preceding claim wherein the sensor means comprise at least one of a load cell (6) and a linear variable displacement transformer (90). A suspension system as claimed in any preceding claim wherein the servo-valve (80) is arranged selectively to apply predetermined positive or negative damping to the hydraulic actuator movement by connection of spaces (70, 74) within the cylinder at opposite sides of the piston to fluid pressure and/or return lines (84,85) so as to apply selected damping to the movement of the piston. A suspension system as claimed in Claim 7 wherein the servo-valve (80) is movable to a position in which both the cylinder spaces (70,74) are in communication with the pressure fluid return line (85), to thereby effect positive damping of the piston mouvement. A suspension system as claimed in Claim 7 or 8 wherein the servo-valve (80) is movable to a position in which the cylinder spaces (70,74) communicate respectively with the pressure fluid supply line (84) and with the return line (85), whereby piston movement in the direction to reduce the volume of the interior communicated with the return line is negatively damped. A suspension system as claimed in any one of the preceding claims wherein one of the cylinder spaces (70,74) is communicated with the return line (85) through a non-return valve. A suspension system as claimed claim 7, 8 or 9 wherein the servo-valve (80) is connected through a pair of fluid lines (88,89) to the first and second spaces and each fluid line is connected via a non-return valve (101, 102) to a return line.

Inventors: WILLIAMS, DAVID ALLEN, WRIGHT, PETER GODFREY
Applicants: GROUP LOTUS PLC
Application year: 1986
Main classification: B60G17
IPC codes: B60G17, F16F9