SPRING-DAMPER SYSTEM

Abstract

A spring-damper system consisting of at Least a differential cylinder (4), a hydraulic accumulator (26) and a control valve device (1, 2), is characterized in that by means of at least one motor-pump unit (22) pressure fluid can be supplied to the annular end (6) or both the annular end (6) and the piston end (8) of the differential cylinder (4) in a dosed circuit using the control valve device (1, 2).

Claims

1. A spring-damper system consisting of at least a differential cylinder (4), a hydraulic accumulator (26) and a control valve device (1, 2), characterized in that by means of at least one motor-pump unit (22) pressure fluid can be supplied to the annular end (6) or both the annular end (6) and the piston end (8) of the differential cylinder (4) in a closed circuit using the control valve device (1, 2).

2. The system according to claim 1, characterized in that the control valve device has two control valves (1, 2), of which in a fluid-conveying manner one control valve (1) is connected at its inlet (10) to the annular end (6) and at its outlet (12) is connected to both the piston end (8) and the inlet (14) of the second control valve (2), the outlet (16) of which is connected to the inlet (18) of the pump (20) of the motor-pump unit (22).

3. The system according to claim 1-er-2, characterized in that the two control valves (1, 2) are proportional throttle valves, preferably electromagnetically actuatable 2/2-way proportional throttle valves.

4. The system according to claim 1, characterized in that the hydraulic accumulator (26) is installed in the connection line (24) between the outlet (16) of the second control valve (2) and the inlet (18) of the pump (20).

5. The system according to claim 1, characterized in that a check valve (34), which opens in the direction of the annular end (6), is installed in the connection line (32) between the outlet (30) of the pump (20) and a branching-off point (38) which is connected to the annular end (6) and to the inlet (10) of the first control valve (1) in a fluid-conveying manner.

6. The system according to claim 1, characterized in that a pressure relief valve (36) is installed between the part of the connection line (32) routed from the outlet (30) of the pump (20) to the check valve (34), and the connection line (24) routed to the pressure accumulator (26).

7. The system according to claim 1, characterized in that the motor-pump unit (22) comprises a gear pump (42), whose leakage oil port (44) is connected to a return line (46).

8. The system according to claim 1, characterized in that the outlet (48) of a feed pump (50) is connected to the inlet (18) of the gear pump (42).

9. The system according to claim 1, characterized in that the motor-pump unit (22) has a radial piston pump (20) or an orbital motor is used instead.

Description

[0015] The invention is explained in detail below with reference to exemplary embodiments shown in the drawing. In the Figures:

[0016] FIG. 1 shows a symbolic representation of the fluid circuit of an exemplary embodiment of the spring-damper system according to the invention;

[0017] FIGS. 2 to 5 show the fluid circuit of FIG. 1, with lines of different line widths indicating four different main states of the system of the exemplary embodiment; and

[0018] FIG. 6 shows a symbolic representation of the fluid circuit of a second exemplary embodiment of the system according to the invention.

[0019] in the figures, a differential cylinder provided as a suspension strut, in particular of a cabin suspension, is designated by the reference numeral 4, has a piston rod designated by 5 and has working chambers of differently effective piston surfaces at its annular end 6 and its piston end 8. The annular end 6 and piston end 8 are connected to a control valve device comprising two control valves, each formed by a proportional throttle valve. The present exemplary embodiments concern electromagnetically controlled 2/2-way proportional throttle valves designated by 1 and 2, respectively. Of these, the proportional throttle valve 1 at its inlet 10 is connected to the annular end 6 of the differential cylinder 4 and at its outlet 12 is connected to both the piston end 8 and to the inlet 14 of the second proportional throttle valve 2. The Latter is connected at its outlet 16 to the inlet 18 of the pump 20 of the motor-pump unit 22 via a connection line 24. The oil end 28 of a hydropneumatic pressure accumulator 26 is also connected to the connection line 24. The outlet 30 at the pressure end of the pump 18 is connected to the annular end 6 of the differential cylinder 4 via a second connection line 32, in which there is a check valve 34 that opens in the direction of the annular end 6. A pressure relief valve 36 interposed between a branch point 38 located at the second connection line 32 between the check valve 34 and the pump outlet 30, and a branch point 40 at the first connection line 24 complements the fluid circuit of the first exemplary embodiment shown in FIGS. 1 to 5.

[0020] In this arrangement, the piston end 8 of the differential cylinder 4, in conjunction with the hydraulic accumulator 26, bears the static load, which can result in a static pressure of more than 100 bar for a standard 3-point support of a cabin weighing 300 kg. In view of the high-pressure level, the pump 20 of the motor-pump unit 22 in this example is an axial piston pump, which permits high pressures at the suction-end inlet 18. Alternatively, an orbital motor could be used.

[0021] As long as the proportional throttle valves 1 and 2 are not actuated and are open in their non-throttling home position, the motor-pump unit 22 does not have to build up any pressure. Apart from the line resistances, the pump 20 pumps the oil without pressure difference in the closed circuit containing the differential cylinder 4, wherein the annular chamber 6 is connected to the pressure-end outlet 30 of the pump 20.

[0022] The piston end 8 is connected to the outlet 12 of the first proportional throttle valve 1 and to the inlet 14 of the second proportional throttle valve 2. As long as both valves 1 and 2 are in their home position, the static pressure at the annular end 6 and at the piston end 8 is identical, and because they are interconnected without throttling, the suspension is undamped. In FIGS. 2 to 5, four main states of the system that occur when the valves 1 and 2 are actuated, are indicated in that the line sections bearing the higher pressure, are drawn using a greater line thickness.

[0023] In the “active compression” state illustrated in FIG. 2, the first proportional throttle valve 1 is actuated from the open home position to move to a throttle position. The volume flow generated by the pump 20 causes a pressure acting in the annular chamber 6 of the differential cylinder 4 to be built up by the throttling effect of the actuated valve 1, wherein said built-up pressure gives rise to an active compression motion of the piston rod 5.

[0024] FIG. 3 refers to the “Active rebound” state. In this state, the second proportional throttle valve 2 is actuated. As a result of its throttling effect, a pressure is built up both in the annular chamber 6 and in the piston chamber 8. Because of the larger piston surface of piston chamber 8, the increased pressure in the cylinder 4 causes an active extending motion of the piston rod 5.

[0025] FIG. 4 refers to the “Damping during rebound” state. In this state, the piston rod 5 of the cylinder 4 performs an extending motion. Oil is thus displaced from the annular chamber 6 to the piston chamber 8 via the open proportional throttle valve 1, which is not actuated. If the proportional throttle valve 1 is now actuated, this volume flow builds up and creates a pressure difference between the annular end 6 and the piston end 8 of the cylinder 4. This pressure difference has a damping effect during the extending motion.

[0026] FIG. 5 refers to the “Damping during compression” state. The piston rod 5 of the cylinder 4 is now in a retracting motion. Some of the oil flows from the piston end 8 to the annular end 6 via the non-actuated, open proportional throttle valve 1, The other part flows into the accumulator 26 via the second proportional throttle valve 2.1f the second proportional throttle valve 2 is now actuated, a pressure difference is built up by the volume flow to the accumulator 26 and counteracts the spring compression in a damping manner.

[0027] FIG. 6 shows a second exemplary embodiment in which the closed circuit having differential cylinder 4, control valves 1 and 2, accumulator 26 and pump of motor-pump unit 22 corresponds to the first exemplary embodiment, The difference, in contrast thereto, is that instead of the axial piston pump 20, a gear pump 42 having a leakage oil port 44 is used and the leakage oil port 44 is connected to a tank 52 via a return line 46 and is thus non-pressurized. The resulting pressure release of the shaft seal of the gear pump 42 renders it pressure-resistant at its two ports and thus safe to operate despite the high pressure Level present in the closed circuit. However, the leakage oil flow, which can be up to 1% of the nominal volume flow, causes the piston rod 5 to continuously subside because of the permanent drain from the closed circuit to the tank 52. To still be able to maintain the desired level position, a feed pump 50 in the form of a small gear pump is provided, which takes in from the tank 52 and generates a feed pressure at its pressure-end outlet 48, which is connected to the connection line 24 via a check valve 54. A proportional pressure relief valve 56, which is installed between the outlet 48 of the feed pump 50 and the tank 52, can be used to adjust the feed pressure and thus the level position. An additional advantage of the embodiment of Fig, 6 is that because of the Leakage and the oil that is permanently re-injected to compensate for the leakage, continuously flushes the closed circuit of the system.