"Vehicle suspension system, especially for use in transport vehicles for open-pit mining with a weight payload greater than 100 tons"

Abstract

Vehicle suspension system for multi-axle wheeled vehicles, comprising a hydraulic line system (4) and a damping system (2) integrated into this line system (4), as well as a pressure accumulator system (1L, 1R, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) connected to this damping system (2) by the line system (4), and spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R) also connected to the line system (4), wherein the damping system (2) has damping valves (15L, 15R) in which an oil chamber (67, 68) is connected to the pressure accumulator system (1L, 1R, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) and an oil chamber (30, 31) is connected to the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R), characterized in that the damping valve (15L, 15R) has a control piston (36, 37) and the control piston (36, 37) has a mushroom-shaped piston part (45, 46) which has a bore (57, 58) into a pot (49,50) can close or open for oil flow, and furthermore, a movable disc (61, 62) is provided in the damping valve (15L, 15R) which can close or open a second bore (59, 60) in the pot (49, 50) for oil flow, and that the mushroom-shaped piston part (45, 46) is arranged in relation to the bore (57, 58) such that, in the event of oil flow through a hydraulic fitting (65, 66) in the damping valve (15L, 15R) towards the oil chamber (30, 31), it closes the bore (57, 58), and in the event of oil flow through a hydraulic fitting (53, 54) in the damping valve (15L, 15R) into the oil chamber (67, 68), this oil flow can lift the mushroom-shaped piston part (45, 46) so that oil flow through the Bore (57, 58) from the oil chamber (67, 68) into the oil chamber (30, 31) becomes possible, and the disc (61, 62) directs the oil flow from the oil chamber (30, 31) to the oil chamber (67, 68) through the bore (59, 58) when oil flows through the hydraulic fitting (65, 66) to the oil chamber (30, 31).60) by lifting the disc (61, 62) and, when the oil flows through the hydraulic fitting (53, 54) into the oil chamber (67, 68), the disc (61, 62) blocks the oil flow through the bore (59, 60), whereby the disc (61, 62) and also the mushroom-shaped piston part (45, 46) are moved by the direction of the oil flow, but in contrast to the disc (61, 62), the mushroom-shaped piston part (45, 46) is additionally controlled in its stroke movement by the control piston (36, 37) and thus a controlled oil throttling is present when the oil flows from the oil chamber (67, 68) to the oil chamber (30, 31) and when the oil flows from the oil chamber (30, 31) to the oil chamber (67, 68) through the disc (61, 62) no intentional Oil throttling is generated, whereby this different flow influence through the mushroom-shaped piston part (45, 46) and the disc (61, 62) creates a damping system which, when the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R) are compressed,During compression, the oil flow from oil chamber (30, 31) to oil chamber (67, 68) does not produce any intended oil throttling, i.e., no damping. However, during expansion of the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R), i.e., during rebound, the oil flow from oil chamber (67, 68) to oil chamber (30, 31) produces controlled oil throttling, i.e., controlled damping.

Description

[0001] The present invention relates to a vehicle suspension system for multi-axle wheeled vehicles, in particular for wheeled vehicles with a weight payload of more than one hundred tons of material to be transported.

[0002] These types of vehicles are used in open-pit mining to transport the extracted material, such as coal, iron ore, copper, etc., along spiral tracks from the mine to an unloading point for further processing. These transport vehicles have a driven front axle and a driven rear axle. To enable the transport of payloads of one hundred tons or more, these vehicles have oversized tires, for example, up to four meters in diameter. These oversized tires are filled with air, and the large air volume provides a cushioning effect. This cushioning effect is used as part of the vehicle's suspension system. Additionally, these vehicles may also have suspension for each wheel, although this suspension has a relatively short travel.This relatively short suspension travel is necessary for stability reasons, as the high center of gravity of such vehicles makes lateral stability very critical, and with greater suspension travel, the lateral tilt angle becomes so large that there is a risk of rollover. Due to this short suspension travel for the vehicle's size, the ride comfort, and therefore the overall comfort, is poor. Another disadvantage is that the tire suspension cannot be damped, causing the vehicle to experience a reciprocating motion while driving. This constant reciprocating motion puts strain on the driver, leading to premature fatigue and consequently to a lack of concentration while driving, which increases the risk of accidents. In addition to the reciprocating motion, the oversized tires also have the disadvantage of being extremely expensive to purchase, for example, up to US$40,000 per tire.In addition, the manufacturing time for such oversized tires is long, and transport from the manufacturing plant to the mines is complex, expensive, and difficult.

[0003] The invention relates to a vehicle suspension system according to claim 1.

[0004] Such a spring system is known from DE 39 36 034 A1, and an end-position damping system is disclosed that decelerates the spring movement of the spring cylinder to zero over a short portion of the compression stroke. However, this end-position damping system is not suitable for implementing displacement-dependent damping of a spring cylinder, since the damping effect of the end-position damping depends solely on the flow velocity of the oil in the throttle gap. Furthermore, the throttle effect of the throttle gap acts as damping only in the compression direction of a spring cylinder. The disclosed end-position damping system is also intended to act only in the compression direction. If it were to act equally in the rebound direction, a spring cylinder would no longer be able to rebound quickly enough, and the wheel would lose contact with the road surface.

[0005] DE 39 36 034 A1 also discloses a hydraulically controlled damping valve.

[0006] The damping valve 46, 48 consists of a hydraulic adjusting unit 48 and an adjustable throttle 46. This damping valve 46, 48 is in Fig. 1 to Fig. Figure 4 of DE 39 36 034 A1 shows that the symbolic representation, which is standard in hydraulics, clearly clarifies that the throttling effect of the throttle 46 should be equally effective in both flow directions through the throttle.

[0007] If a differential throttling dependent on the direction of oil flow had been desired, two throttles 46 would have been provided. Additionally, flow-dependent shut-off valves would have been necessary to control the oil flow through one throttle 48 during compression and through another throttle 48 during rebound. However, a damping valve 46, 48 with only one throttle 48 is provided, thus ensuring identical oil throttling in both flow directions.

[0008] The disadvantage of this equal oil flow restriction in both directions is that the damping of the wheel's stroke is the same in both compression and rebound directions. For a comfortable ride, however, less restriction of the stroke in the compression direction is advantageous, while when cornering, greater restriction of the oil flow in the rebound direction is beneficial to reduce body roll.

[0009] This damping valve has the crucial disadvantage that, when the throttle 46 is hydraulically adjusted, it alters the throttling effect in both the compression and rebound directions. This damping system does not allow the damping effect to be controlled differently in the compression direction by means of an oil pressure control of the throttle than in the rebound direction. Furthermore, it is not possible with the throttle valve 46, 48 to have no throttling effect in the compression direction and simultaneously generate appropriately controlled, travel- and load-dependent damping in the rebound direction.

[0010] With a vehicle suspension, you feel road irregularities if the suspension is stiff, as the compression of the suspension causes shocks to the vehicle. These shocks also cause damage to the road surface, as the impact is transmitted to the road itself.

[0011] For ride comfort and to protect the road surface, the vehicle's spring rate should be as soft as possible in the compression direction. However, a soft suspension results in the vehicle leaning significantly to the side when cornering, leading to unpleasant and unsafe handling. This drastically reduces ride comfort.

[0012] In the direction of rebound, the wheel must move quickly, otherwise it will lose contact with the road surface. However, it must not move too quickly in the direction of rebound, as this could cause it to strike the road and create a shock. Additionally, sufficient rebound travel is necessary to prevent potholes in the road surface from causing the wheel to lose contact. However, large rebound travel has the disadvantage of increasing the vehicle's tendency to roll.

[0013] For driving safety, the vehicle suspension should also adjust its behavior to different vehicle loads, since the wheel's lifting motion must be damped differently when the vehicle is loaded compared to when it is empty. Lateral roll during cornering or lateral movement also differs between a loaded and an empty vehicle. The throttle system with damping valve 46, 48 shown in DE 39 36 034 therefore cannot meet the requirements for a very comfortable and road-friendly vehicle suspension.

[0014] The present invention is based on the objective of creating a vehicle suspension system for wheeled vehicles with a payload of more than one hundred tons, which can cushion vehicle axles with large suspension travel and, despite the large suspension travel, has such high lateral stability that no critical driving condition arises due to the high center of gravity and the large suspension travel, and the contribution of the tire suspension to the vehicle suspension system is such that vehicle tires without air filling and thus without lifting oscillation can be used.

[0015] According to the invention, this is achieved by the features of claim 1. The dependent claims contain advantageous features of the vehicle spring system according to the invention, which, individually or in combination with one another, represent a further embodiment of the subject matter of claim 1.

[0016] In the vehicle suspension system according to the invention, the vehicle wheels, coated with a relatively thin, wear-resistant tread layer, are sprung by spring cylinders, with the vehicle wheels being guided by swing arms. The compression movement of the spring cylinders is not influenced by a damping unit; however, in the rebound direction, the movement of the spring cylinders of the vehicle suspension system is damped by an oil-pressure-controlled damping unit, depending on the load and spring travel. The very long total spring travel of the vehicle suspension system according to the invention is such that the spring travel in the rebound direction is significantly longer than that in the compression direction, preferably approximately two-thirds in the rebound direction and one-third in the compression direction.The spring force of the spring cylinders is generated with the help of a nitrogen pressure storage system, whereby the pressure of the nitrogen in the storage system is transferred to the spring cylinders by means of oil guided in pipes.

[0017] The invention will be explained in more detail with reference to the exemplary embodiments shown in the figures. The figures show: Fig. 1 Schematic view of a vehicle spring system according to the invention for an axle line with oil pressure-dependent damping control, Fig. 2. Schematic view of a vehicle spring system according to the invention for an axle line with oil pressure-dependent damping control and weight load compensation between the right and left sides of the vehicle. Fig. 3 Schematic view of a vehicle suspension system according to the invention for two axle lines with oil pressure-dependent damping control and weight load compensation between the right and left sides of the vehicle, Fig. 4 Schematic view of a vehicle suspension system according to the invention for two axle lines with oil pressure-dependent damping control, weight load compensation between the right and left sides of the vehicle and a pressure storage system consisting of three pressure accumulators, Fig. 5 Schematic view of a vehicle spring system according to the invention for three axle lines with oil pressure-dependent damping control, weight load compensation between the right and left sides of the vehicle and a pressure storage system consisting of five pressure accumulators, Fig. 6 Sectional drawing of a damping valve according to the invention for pressure-controlled damping in the rebound direction, Fig. 7 Sectional drawing of a damping valve according to the invention for pressure-controlled damping in the extension direction with control of the damping by the spring movement of the opposite side of the vehicle, Fig. 8 Sectional drawing of a detail of a damping valve according to the invention with pressure-controlled rebound damping in the closed position, Fig. 9 Sectional drawing of a detail of a damping valve according to the invention with pressure-controlled rebound damping in the open position, Fig. 10 Control pistons of a damping valve according to the invention with circumferential thickening in the central region of the control piston, Fig. 11 Sectional drawing of a damping valve according to the invention, Fig. 12 to 15 Sectional drawing of a detail of a damping valve according to the invention, Fig. 16 Schematic view of a vehicle spring system according to the invention with sectional drawing of the spring cylinders, Fig. 17 to 22 Sectional drawing of a damping unit according to the invention with oil system pressures, Fig. 23 to 25 Schematic view of a flow distributor according to the invention, Fig. 26 and Fig. 27 Sectional drawing of a pressure storage system according to the invention, Fig. 28 Schematic diagram of a spring curve, Fig. 29 and Fig. 30 Sectional drawing of a pressure storage system according to the invention, Fig. 31 and Fig. 32 Schematic drawing of a vehicle wheel according to the invention with elastic coating of the tread, Fig. 33 and Fig. 34 Schematic drawing of a holding device according to the invention for each vehicle wheel, Fig. 35 Schematic view of a vehicle suspension system according to the invention for two axle lines with hydraulic lifting and lowering device for the vehicle.

[0018] In the various figures of the drawing, identical parts are always marked with the same reference symbols. For clarity, parts assigned to the left side of the vehicle are marked with an additional L, and parts for the right side with an additional R.

[0019] In the Fig. Figures 1 to 5 show advantageous embodiments of a vehicle spring system according to the invention. A vehicle suspension system, particularly for wheeled vehicles with a payload exceeding one hundred tons, used as transport vehicles in open-pit mining, comprises a pressure accumulator system 1L, 1R, a pressure-controlled damping system 2, spring cylinders 3L, 3R, 3.1L, 3.2L, 3.3L, 3.1R, 3.2R, 3.3R, a hydraulic line system 4, and vehicle wheels 6L, 6R, 6.1L, 6.2L, 6.3L, 6.1R, 6.2R, 6.3R with an elastic coating 5 on their outer surface, and these are rotatably mounted in wheel-guiding swing arms 9L, 9R in a bearing 8L, 8R, 8.1L, 8.2L, 8.3L, 8.1R, 8.2R, 8.3R in the vehicle frame 7. 9.1L, 9.2L, 9.3L, 9.1R, 9.2R, 9.3R stored.

[0020] As in Fig. 1 to Fig. As shown in Figure 5, one pressure storage system 1L acts on the spring cylinders 3L, 3.1L, 3.2L, 3.3L via the line system 4 and the damping system 2. The other pressure storage system 1R acts on the spring cylinders 3R, 3.1R, 3.2R, 3.3R via the line system 4 and the damping unit 2. Based on the understanding that comfortable vehicle suspension can only be achieved if no or minimal damping force acts on the stroke of the wheel 6L, 6R, 6.1L, 6.2L, 6.3L, 6.1R, 6.2R, 6.3R in the compression direction, the line system 4 is designed such that several hydraulic lines 4 are combined via flow distributors 11.1L, 11.2L, 11.1R, 11.2R, 12L, 12R, 13.1L, 13.2L, 13.1R, 13.2R, 14L, 14R, which are described in more detail below. Fig. 3 to Fig. 5 and Fig. 35, and not, as is usually the case with hydraulic piping, with T-fittings.

[0021] As in Fig. 1 to Fig. As shown in Figure 7, the pressure-controlled damping system comprises two damping valves 15L, 15.1L, 15R, and 15.1R, a valve unit 16 with double check valves 17 and 18, pressure accumulators 19 and 20, and hydraulic lines 21, 22, 23, 24, 25, 26, 27, 28R, 28L, 29, 34, 35, and 149. The double check valves 17 and 18 are briefly opened by the oil pressure of an oil pump 149, by applying oil pressure to the double check valves 17 and 18 via a line 150. With the double check valves 17, 18 open, oil chambers 30, 31 are connected to oil chambers 32, 33 via lines 21, 22, 23, 24. Simultaneously, the pressure accumulator 19, 20 is connected to oil chambers 30, 31, 32, 33 via oil lines 34, 35. When line 150 is depressurized, the double check valves 17, 18 close, and the oil pressure present in oil chambers 30, 31 at that moment is stored in the pressure accumulator 19, 20 and oil chambers 32, 33.The oil pressure prevailing in oil chamber 30 at the time the double check valve 17 is open is stored in oil chamber 32 and pressure accumulator 19, while the oil pressure prevailing in oil chamber 31 is stored in oil chamber 33 and pressure accumulator 20. Pressure accumulators 19 and 20 ensure that the incompressible oil volume contained in oil chambers 32 and 33 does not impede the movement of a control piston 36, 37, 38, or 39. Pressure accumulators 19 and 20 are dimensioned so that a stroke of the control pistons 36, 37, 38, or 39 does not result in any control-influencing change in the oil pressure in oil chambers 32 and 33.

[0022] As in Fig. 6, Fig. 8 and Fig. As shown in Figure 9, the control piston 36, 37 has a seal 43, 44 on its outer surface 40, 41, which seals the oil chamber 32, 33 against the oil chamber 30, 31. On the side of the control piston 36, 37 opposite the oil chamber 32, 33, there is a mushroom-shaped piston part 45, 46, which rests on a base 47, 48 of a pot 49, 50 with a bore 57, 58 and closes the bore 57, 58, see Figure 9. Fig. 8, or not resting, leaving bore 57, 58 open, see Fig. 9.

[0023] The pot 49, 50 is clamped against a shoulder 55, 56 via a bushing 51, 52 with a hydraulic screw connection 53, 54.

[0024] In the base 47, 48 of the pot 49, 50 there is a second bore 59, 60, which can be closed by a disc 61, 62, see Fig. 9, or the disc 61, 62 does not close the bore 59, 60, see Fig. 8. The disk 61, 62 is mounted in a basket 63, 64 connected to the underside of the base 47, 48, in which bores 63.1, 63.2 are located.

[0025] The disc 61, 62 rests against the underside of the pot 49, 50 and closes the bore 59, 60, see Fig. 9, when the oil flows from the pressure accumulator system 1L, 1R via the hydraulic fitting 53, 54 towards the oil chamber 30, 31 during a rebound movement of the spring cylinders. The disc 61, 62 lifts off the bore 59, 60, see Fig. 8, when the oil flows from there via the line system 4 and a hydraulic fitting 65, 66, 75, 76 into the oil chamber 30, 31 during the compression movement of the spring cylinders 3L, 3R, 3.1L, 3.2L, 3.3L, 3.3R and further towards the pressure accumulator system 1L, 1R.

[0026] When the spring cylinders 3L, 3R, 3.1L, 3.2L, 3.3L, 3.1R, 3.2R, 3.3R extend, oil flows from oil chamber 30, 31 through the hydraulic fitting 65, 66 and through the line system 4 towards the spring cylinder, oil is drawn from oil chamber 67, 68 through bore 59, 60. This suction action causes the disc 61, 62 to be drawn against the underside of the pot 49, 50, closing bore 59, 60. The oil pressure in oil chamber 30, 31 decreases. A pressure differential arises between the oil pressure in chambers 30, 31 and the oil pressure of the storage system 1L, 1R, such that the oil pressure in the storage system is higher than that in chambers 30, 31, and therefore the oil pressure in the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R is lower than in the storage system 1L, 1R. As already described above, the oil pressure in oil chambers 32, 33 causes the control piston 36, 37 with the mushroom-shaped piston part 45, 46 to be pressed against the base 47, 48, thus closing the bore 57, 58.The oil pressure differential between oil chambers 67, 68 and 30, 31 increases until the higher oil pressure in oil chamber 67, 68 moves the control piston 36, 37 against the oil pressure in oil chamber 32, 33. From this point on, the oil can flow from the pressure accumulator system 1L, 1R via the damping valve 15L, 16R to the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R. The oil pressure in oil chamber 32, 33 ensures that during the entire extension movement of the spring cylinder, the control pistons 36, 37 are always pressed with force towards the ground 47, 48, thus maintaining a pressure differential between oil chamber 67, 68 and oil chamber 30, 31 throughout the entire extension travel. This.

[0027] The pressure differential causes the spring force of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R to decrease during rebound. The difference in force between the spring cylinders during rebound and at rest is the damping of the vehicle's suspension system.

[0028] Since the oil pressure in oil chambers 30, 31 corresponds to the oil pressure in spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R, and the oil pressure in the spring cylinder in turn corresponds to the vehicle load, the oil pressure in oil chambers 30, 31 depends on the vehicle's load. As described above, opening the hydraulic check valves 17, 18 directs the oil pressure from oil chambers 30, 31 into oil chambers 32, 33. Consequently, the control pistons 36, 37 are pushed towards the ground 47, 48 according to the vehicle load, and thus the vehicle's suspension system is damped in the rebound direction according to the vehicle load.

[0029] How Fig. Figure 16 shows that the spring cylinder 3L, 3R {3.1L, 3.2L, 3.3L, 3.1R, 3.2R, 3.3R} has a piston 108, 109 with seal 110, 111 and the piston rod 108.1, 109.1 attached to the piston 108, 109. The seal 110, 111 divides the cylinder chamber into a piston chamber 112L, 112R and an annular chamber 113L, 113R.

[0030] The piston chamber 112L, 112R of the spring cylinder 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R is connected to the oil chamber 30, 31 via a hydraulic line 69L, 69R through the hydraulic fitting 65, 66, see Fig. 1. The annular space 113L, 113R is connected to the pressure accumulator system 1L, 1R via a hydraulic line 70, 71.

[0031] In the piston chamber 112L, 112R of the spring cylinder 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R, the reduced oil pressure due to the damping valve 15L, 15R acts, while in the annular chamber 113L, 113R, the higher pressure of the accumulator system 1L, 1R acts. During the rebound movement, the load-bearing capacity of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R is further reduced by the counter-pressure of the accumulator system 1L, 1R prevailing in the annular chamber 113L, 113R, since the load-bearing capacity of the spring cylinder is equal to the piston area times the oil pressure in the piston chamber minus the annular chamber area times the oil pressure in the annular chamber. The counter-pressure of the storage system 1L, 1R in the annular space 113L, 113R of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R acts as an amplifier of the rebound damping. This amplification does not act in the compression direction, since, according to the invention, the compression movement is not damped by damping elements and thus, during the compression movement of the spring cylinders, there is no change in the oil pressure in the spring cylinders 3L, 3R, 3.1L, 3.2L, 3.3L, 3.3R.1R, 3.2R, 3.3R and the oil pressure in the pressure accumulator system 1L, 1R, there is no pressure difference generated by controlled damping elements in the damping valve 15L, 15R, 15.1L, 15.1R.

[0032] An unavoidable pressure loss during spring cylinder movement is minimized by designing the piping system 4 to achieve the lowest possible flow losses in the compression direction of the spring cylinders. Similarly, the piping system 4 is optimized in the rebound direction to minimize flow losses, ensuring that the oil volume flowing into the annular space 113L, 113R during compression is not dampened by throttling losses. Damping the oil volume flowing into the annular space 113L, 113R would lead to a pressure drop within the space. As already mentioned above, such a pressure reduction in the annular space 113L, 113R increases the spring force (load-bearing capacity) of the spring cylinder 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R and is noticeable during the suspension movement as a damping force.

[0033] An advantageous measure for reducing pressure loss in the compression and rebound directions of the spring cylinders are the flow distributors 11.1L, 11.2L, 11.1R, 11.2R, 12L, 12R, 13.1R, 13.2R, 13.1L, 13.2L, 14L, 14R, 15L, 15.1L, 15R, 15.1R, installed in the line system 4 according to the invention. Fig. 3 to Fig. 5.

[0034] From the finding that a compression movement only triggers a rolling movement around the longitudinal axis of the vehicle on one side, we show Fig. 2 and Fig. 7 an advantageous embodiment of the vehicle suspension system that suppresses the rolling motion about the longitudinal axis of the vehicle better than the one in Fig. 1 shown.

[0035] In order for the damping valve 15.1L, 15.1R of the spring system to recognize that a rolling movement of the vehicle can be triggered, the compression movement of the spring cylinders on one side of the vehicle is used, according to the invention, by a hydraulic and not an electrical signal for damping control of the damping valve on the other side of the vehicle.

[0036] According to the invention, this damping control does not act on the throttling of the oil flow in the compression direction, but only on the throttling of the oil flow in the rebound direction of the spring cylinders. Thus, according to the invention, only the damping of the rebound movement of the spring cylinders is used to suppress the lateral roll movement of the vehicle, and not damping of the spring cylinders in the compression direction.

[0037] The piston chamber 112L, 112R of the spring cylinders 3L, 3R, {3.1L, 3.2L, 3.3L, 3.1R, 3.2R, 3.3R} is connected to the oil chamber 30, 31 of the damping valve 15L, 15R, 15.1L, 15.1R via the hydraulic line 69L, 69R and the hydraulic fitting 65, 66, see Fig. 2 and Fig. 16. The annular space 113L of the spring cylinders 3L, {3.1L, 3.2L, 3.3L} is connected to the oil chamber 31 in the damping valve 15R, 15.1R via a hydraulic line 72 and a hydraulic fitting 74, and the annular space 113R of the spring cylinders 3R, {3.1R, 3.2R, 3.3R} is connected to the oil chamber 30 in the damping valve 15L, 15.1L via a hydraulic line 73 and a hydraulic fitting 75.

[0038] The oil pressure in the annular space 113L of cylinder 3L, {3.1L, 3.2L, 3.3L} is therefore the same as in the piston chamber 112R of cylinder 3R, {3.1R, 3.2R, 3.3R}, since both cylinder chambers 113L, 112R are connected to the chamber 31 in the damping valve 15R, 15.1R, and the oil pressure in the annular space 113R of cylinder 3R, {3.1R, 3.2R, 3.3R} is the same as in the piston chamber 112L of cylinder 3L, {3.1L, 3.2L, 3.3L}, since both cylinder chambers are connected to the chamber 30 in the damping valve 15L, 15.1L. This interconnection of the spring cylinder chambers 112L, 113R, 112R, 113L increases the damping of the spring cylinders in the rebound direction on the respective other side of the vehicle.

[0039] As weight measurements have shown, it is normal for the weight of one side of a wheeled vehicle to differ from the weight of the other. Due to these weight differences, the spring load on the two sides can vary significantly without additional measures. The connection between piston chamber 112L, 112R and the annular chamber 113L, 113R on the other side of the vehicle not only supports damping in the rebound direction, as described above, but, more importantly, improves ride comfort by balancing the spring load between the two sides, as demonstrated in the following example calculation.

[0040] The annular chamber 113L is hydraulically connected to the piston chamber 112R, and thus both chambers are hydraulically synchronized and the annular chamber 113R is hydraulically synchronized with the piston chamber 112L.

[0041] The load-bearing capacity of a hydraulic cylinder is the oil pressure on the area of ​​the piston (108, 109) minus the oil pressure on the area of ​​the annular space (113L, 113R). Using the Fig. Section 16 explains the load-balancing effect of this hydraulic circuit.

[0042] The weight load G1 acts on the spring cylinder 3L, {3.1L, 3.2L, 3.3L} on the left side of the vehicle, and the weight load G2 acts on the spring cylinder 3R, {3.1R, 3.2R, 3.3R} on the right side of the vehicle. The weight load G1 is assumed to be 30 tons and the weight load G2 is assumed to be 40 tons. In a preferred embodiment for such a load, the spring cylinder 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R has a piston diameter 108, 109 of 255 mm and a piston rod diameter 108.1, 109.1 of 200 mm. With these dimensions, a hydraulic pressure of 104.4 bar is established in piston chamber 112L for supporting the 30 tons with cylinder 3L, {3.1L, 3.2L, 3.3L}, and an oil pressure of 118.5 bar is established in piston chamber 112R by spring cylinder 3R, {3.1R, 3.2R, 3.3R}. Without the line connection 72, 73 and instead the line connection 70, 71, as shown in Fig. As shown in Figure 1, an oil pressure of 95.5 bar is established in piston chamber 112L and an oil pressure of 127.5 bar in piston chamber 112R.

[0043] The circuit with line connection 72, 73, as in Fig. 2 to 5 and Fig. Figure 16 shows that it reduces the oil pressure difference between the right and left sides of the vehicle in the case of asymmetrical weight distribution and, as shown in the example calculation, reduces the pressure difference from 31.9 bar to 14.1 bar.

[0044] For the vehicle suspension and also for the damping system 2, this means that there is a more balanced suspension and damping behavior between the left and right sides of the vehicle.

[0045] To ensure that the compression movement of the other side of the vehicle controls the damping of the rebound movement, the weight load of the other side of the vehicle, in the form of the oil pressure of that side, is transferred from oil chamber 30 through hydraulic lines 22 and 21 into oil chamber 32, and the oil pressure of oil chamber 31 is transferred through hydraulic lines 24 and 23 into oil chamber 33. This is controlled, as already described for the damping valves 15L and 15R, by briefly opening the double check valves 17 and 18 using hydraulic pressure from the oil pump 149. Therefore, after the double check valves 17 and 18 close, the static oil pressure of oil chamber 30 is stored in the pressure accumulator 19, and the static oil pressure of oil chamber 31 is stored in the pressure accumulator 20.

[0046] As described for damping valve 15L, 15R, the rebound damping is controlled by means of the control piston 36, 37. The control piston 38, 39 in damping valve 15.1L, 15.1R has the same control function, see Fig. 7. The control piston 38 has a circumferential enlargement 76 in the center and the control piston 39 has a circumferential enlargement 77 such that the control piston 38 has two additional piston surfaces 78, 79 and the control piston 39 has two additional surfaces 80, 81, see Fig. 7, Fig. 10 and Fig. 11.

[0047] On the control piston 38, 39 are three seals 82, 83, 84, 85, 86, 87, which seal the oil chambers 88, 89, 90, 91. On the control piston 38, seal 82 seals between oil chamber 30 and oil chamber 88, seal 84 seals between oil chamber 32 and oil chamber 90, and seal 83 seals between oil chambers 88 and 90. On the control piston 39, seal 85 seals oil chamber 31 against oil chamber 89, seal 87 seals oil chamber 33 against oil chamber 91, and seal 86 seals between oil chambers 89 and 91.

[0048] To prevent a higher oil pressure in oil chamber 30 or 31 on one side of the vehicle from leading to an unwanted increase in rebound damping of the less loaded suspension on the other side of the vehicle, oil chamber 32 is connected via a line 28R to oil chamber 88 in damping valve 15.1R and oil chamber 33 is connected via a line 28L to oil chamber 89 in damping valve 15.1L, see Fig. 7.

[0049] The oil pressure in oil chamber 32, 33 acts on a piston area 92, 93 of the piston 38, 39 and, via the oil line 28R, 28L, also in oil chamber 88, 89 and thus on the piston area 78, 80. The piston area 78, 80 on the circumferential enlargement 76, 77 of the control piston 38, 39 is dimensioned relative to the piston area 92, 93 such that, in a favorable design, the piston area 78, 79, 80, 81 is approximately twice as large as the piston area 92, 93.

[0050] As mentioned above, a compression movement of the spring cylinders on one side of the vehicle adjusts the damping in the rebound direction of the spring cylinders on the other side. Since, during a compression movement, oil is forced from spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R through the line system 4 and damping system 2 into the pressure accumulator system 1L, 1R, the oil pressure in the pressure accumulator system increases, and consequently also in oil chambers 30, 31. This pressure change in oil chambers 30, 31 is used to control the damping adjustment of the damping valves 15L, 15R, 15.1L, 15.1R.

[0051] Oil chamber 30 is connected to oil chamber 91 via hydraulic line 27, and oil chamber 31 is connected to oil chamber 90 via hydraulic line 29, see Fig. 7 and Fig. 11. The oil pressure of oil chamber 31 acts on the piston area 79, which is the same size as the opposite piston area 78, and the oil pressure of oil chamber 30 acts on the area 81, which is the same size and opposite the piston area 80. The areas 78, 80 are opposite the areas 79, 81 and thus form a stroke adjustment system for control pistons 38, 39 which moves the control piston 38, 39 towards the bore 57, 58 when there is a compression movement of the spring cylinders on the opposite side of the vehicle from the driving position and moves the control piston 38, 39 in the opposite direction when there is a rebound movement of the spring cylinders on the other side of the vehicle from the driving position.

[0052] In order to ensure that, during static spring movement, the control piston 38, 39 is controlled in such a way that a mushroom-shaped piston part 94, 95 does not obstruct the oil flow through the bore 57, 58, a piston surface 96, 97 is positioned opposite the piston surface 92, 93, see Fig. 12 to Fig. 15. The opposite piston surface 96, 97 is created by the fact that the mushroom-shaped piston part 94, 95 is not rigidly connected to the control piston 38, 39, but rather the control piston 38, 39 has a step 98, 99 and a conical piston part 100, 101, and the mushroom-shaped piston part 94, 95 has a stepped bore 102, 103 in which the control piston 38, 39 is displaceable, see Fig. 12 and Fig. 13. Further explanation of the stepped boreholes 102 and 103 follows below.

[0053] The stepped bore 102, 103 is connected to the oil chamber 30, 31 by bores 104, 105 such that when the conical piston part 100, 101 rests against a step 106, 107 in the stepped bore 102, 103, there is no oil connection from the oil chamber 30, 31 to the oil chamber 67, 68. If the conical piston part 100, 101 does not rest against the step 106, 107, there is an oil connection from the oil chamber 30, 31 to the oil chamber 67, 68.

[0054] The conical piston section 100, 101 and the ring area 98, 99 on the piston 38, 39 together equal the area of ​​the opposite piston 92, 93. The oil pressure of the oil chamber 30, 31 thus acts on an area of ​​the same size as the oil pressure in the oil chamber 32, 33.

[0055] Each control piston 38, 39, together with the mushroom-shaped piston section 94, 95, is subjected to the oil pressure of the suspension system on the same side of the vehicle and also to the oil pressure of the suspension system on the other side of the vehicle. Since these oil pressures act on surfaces, the area times the oil pressure constitutes the actuating force for the damping adjustment, and simultaneously the oil pressure is the signal for the necessary adjustment or non-adjustment of the control piston 38, 39 in the damping valve 15.1L, 15.1R.

[0056] Fig. Figure 17 shows the oil pressures present in damping system 2, which are present, for example, due to the asymmetrically loaded suspension in damping system 2. The oil pressures calculated from the exemplary asymmetric load are for stationary vehicle motion, i.e., no extension or compression of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R.

[0057] On the control piston 38 of the damping valve 15.1L, the oil pressure stored in the pressure accumulator 20 (118.5 bar) and the oil pressure of the oil chamber 31 (118.5 bar) are opposite each other at the circumferential enlargement 76 with surfaces 78, 79. Similarly, on both surfaces 80, 81 of the control piston 39 of the damping valve 15.1R, the oil pressure stored in the pressure accumulator 19 (104.4 bar) and the oil pressure of the oil chamber 30 (104.4 bar) are opposite each other. Since the same oil pressures act on the opposing surfaces of the control pistons 38 and 39, no adjusting forces are exerted on the pistons. Surface 92 on the control piston 38 is opposite surface 96 and both surfaces are subjected to an oil pressure of 104.4 bar, and surface 93 on the control piston 39 is opposite surface 97 and both surfaces are subjected to an oil pressure of 118.5 bar.This application of oil pressure to surfaces 92 and 96 on the control piston 38 and to surfaces 93 and 99 on the control piston 39 does not cause any adjustment of the control pistons 38, 39, since the same pressures act on the opposite surfaces and thus the same forces are present.

[0058] The mushroom-shaped piston part 94, 95 is subjected to the same oil pressure of 104.4 bar and 118.5 bar, respectively, on both sides. When the mushroom-shaped piston part 94, 95 rests against the surface 45, 46, the area exposed to the oil pressure of the oil chamber 30, 31 is larger than the side exposed to the oil pressure of the oil chamber 67, 68, since, in order to seal the bore 57, 58, the diameter of the bore 57, 58 is smaller than the outer diameter of the mushroom-shaped piston part 94, 95 (see figure). Fig. 12 and Fig. 13. This results in the surfaces of the piston part 94, 95, which are subjected to pressure on both sides, being of different sizes. The mushroom-shaped piston part 94, 95 is pressed against the base 45, 46 by the force generated by the difference in surface area.

[0059] The same holding effect is also achieved when the disc 61, 62 is in contact with the underside of the base 45, 46, since, in order to be able to seal the bore 59, 60, the outer diameter of the disc 61, 62 is larger than the diameter of the bore 59, 60.

[0060] As described above, during a compression movement of the spring cylinders, the oil in the damping valve 15L, 15R, 15.1L, 15.1R flows through the bore 59, 60 to the pressure accumulator system 1L, 1R. In order to overcome the holding force on the disc 61, 62, the oil pressure in the oil chamber 30, 31 must be higher than the oil pressure in the oil chamber 67, 68. This steeply rising oil pressure, generated by the compression movement, would manifest as a shock to the spring cylinders, negatively impacting suspension comfort. According to the invention, this shock is minimized by shifting the control piston 38, 39 towards the oil chamber 32, 33 as soon as the pressure on the piston surfaces of the control piston 38, 39 is equal, as explained below. This displacement causes the conical piston part 100, 101 to lift off from the stage 106, 107, and oil can flow through the stepped bore 102, 103 into the oil chamber 67, 68.A small amount of oil flows through bores 102 and 103; however, the amount flowing through these bores is sufficient to reduce the oil pressure peak. The increase in oil pressure in oil chambers 30 and 31 is gentler, and the flow force of the oil allows the disk 61 and 62 to lift off the underside of the base 45 and 46 almost gently and without sudden force, and the oil has free flow from oil chambers 30 and 31 to oil chambers 67 and 68.

[0061] Essential for the control of the control piston 38, 39 is that the movement of the control piston 38, 39 is strongly damped, as otherwise uncontrolled oscillation of the control piston would occur. According to the invention, the strong damping of the control piston is achieved by correspondingly large seals 82, 83, 84, 85, 86, 87, which press against the walls of the cylinder bores with a greater preload than is necessary for the actual seal.

[0062] The frictional force of the sealing damping of the control piston 38, 39 is advantageously between 10 kp and 15 kp. A further property desired according to the invention is that, due to the sealing damping of the control piston 38, 39, it is in a position during static spring movement that allows a small oil flow through the bore 102, 103 between the oil chambers 30, 31 and the oil chambers 67, 68.

[0063] According to the invention, the adjustment stroke of the control piston 38, 39 is limited between 0.5 mm and 1.5 mm; advantageously, the adjustment stroke is 1.0 mm. Since the seals 82, 83, 84, 85, 86, 87 deform by approximately 0.5 mm when the pistons 38, 39 are moved, it is ensured that the stepped bore 102, 103 always has flow due to the seal return when no adjustment force acts on the control piston 38, 39 in the direction of the stepped bore 102, 103.

[0064] Since the seals 82, 83, 84, 85, 86, 87 deform elastically during the movement of the control piston 38, 39, this deformation has a restoring effect on the piston 38, 39 towards the oil chamber 32, 33. When the control piston 38, 39 is pressed into the mushroom-shaped piston part 94, 95 by the oil pressure differential, and subsequently the pressure differential is such that there is a force equilibrium on the control piston 38, 39, the restoring force of the seal deformation moves the control piston approximately 0.5 mm towards the oil chamber 32, 33, thereby slightly opening the stepped bore 102, 103 and thus establishing an oil connection from oil chamber 67, 68 via bore 102, 103 and bore 104, 105 to oil chamber 30, 31, thus creating this opening of the The bore 102, 103, 104, 105 is sufficient to minimize or avoid a pressure shock when the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R begin to compress.

[0065] During a uniform compression movement of both spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R from the driving position, a so-called parallel suspension, the oil pressure increases on both sides of the vehicle suspension. Thus, depending on the pressure increase in the right or left vehicle suspension system, both control pistons 38, 39 in the damping valves 15.1L, 15.1R are moved either towards the oil chambers 67, 68 or towards the oil chambers 32, 33, as described above. Fig. 18 and the example calculation below will illustrate this.

[0066] The oil pressure in the spring system increases, for example, from the static system pressure in oil chamber 30 of 104.4 bar to 188 bar, and from the static system pressure in oil chamber 31 of 118.5 bar to 213 bar. The control piston 38 in the damping valve 15.1L is moved towards oil chamber 67, and the control piston 39 in the damping valve 15.1R is moved towards oil chamber 68, as shown by the following comparative calculation.

[0067] The piston areas 92, 93 and 96, 97 of the control piston 38, 39 are represented with a unit of measurement of "1". The piston areas 78, 79 and 80, 81 of the control piston 38, 39 are represented with a unit of measurement of "2", since, as written above, the areas 78, 79, 80, 81 should advantageously be twice as large as the areas 92, 93, 96, 97.

[0068] For the comparative calculation, a positive force applies if it points in the direction of oil chambers 67, 68 and a negative force if it points in the direction of oil chambers 32, 33.

[0069] The oil pressures and thus the force of the adjustment act on the control piston 38 [(213 bar - 118.5 bar) × 2] + [(104.4 bar - 188 bar) × 1] = +189 bar + -83.6 bar = +105.4 bar.

[0070] The control piston 38 is pressed towards oil chamber 67 with 105.4 bar times the area unit “1”.

[0071] The oil pressures and thus the force of adjustment act on the control piston 39 [(188 bar - 104.4 bar) × 2] + [(118.5 bar - 213 bar) × 1] = +167.2 bar + -94.5 bar = +72.7 bar.

[0072] The control piston 39 is pressed towards oil chamber 68 with 72.7 bar times the area unit “1”.

[0073] Despite the adjustment of the control pistons 38 and 39 towards the oil chambers 67 and 68, and thus the closure of bores 57 and 58 and the stepped bores 102 and 103, the oil can flow without damping through bores 59 and 60 from oil chambers 30 and 31 to oil chambers 67 and 68, and from there into the pressure accumulator system 1L and 1R. Therefore, when the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, and 3.3R are in a parallel compression direction, the damping direction is not damped by the damping valves 15.1L and 15.1R.

[0074] When the parallel-sprung spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R retract towards the driving position, oil flows from the pressure accumulator system 1L, 1R through the damping system 2 to the spring cylinders. Due to the flow direction from oil chamber 67, 68 through bore 59, 60 to oil chamber 30, 31, the disc 61, 62 is pressed against bore 59, 60. The oil flow through this bore is blocked. As shown in the calculation above, the control pistons 38, 39 are pressed towards the oil chamber 67, 68, thus blocking the flow of oil through the bore 57, 58 and consequently also through the stepped bore 102, 103. Since no oil can flow from the oil chamber 67, 68 into the oil chamber 30, 31, the oil pressure in the oil chamber 30, 31 drops, and consequently also in the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R, and continues to do so until the higher oil pressure in oil chamber 67, 68 pushes the mushroom-shaped piston 94, 95 together with the control piston 38, 39 towards oil chamber 32, 33. From this point on, the oil can flow through the gap between the mushroom-shaped piston 94, 95, which has been lifted from the base 47, 48. Despite the oil flow through bore 57, 58, an oil pressure differential remains between oil chamber 67, 68 and 30, 31, because the control piston 38, 39 is still being pushed towards oil chamber 67, 68. The pressure drop between oil chambers 67, 68 and oil chambers 30, 31, and also in spring cylinders 3L, 3R, depends on the force with which the control pistons 38, 39 are pressed towards oil chambers 67, 68. The actuating force on the control pistons 38, 39 towards oil chambers 67, 68 decreases with the extension travel towards the driving position. In the driving position, the actuating force is zero due to the oil pressure difference. Spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R extends further beyond the driving position with parallel suspension, thus increasing the actuating force on the control pistons 38, 39 again, as with . Fig. shown in 19 and the following invoice.

[0075] The oil pressures and thus the force of adjustment act on the control piston 38 [(90 bar - 118.5 bar) × 2] + [(104.4 bar - 80 bar) × 1] = -57 bar + 24.4 bar = -32.6 bar.

[0076] The control piston 38 is pressed towards oil chamber 32 with 32.6 bar times the area unit “1”.

[0077] The oil pressures and thus the force of adjustment act on the control piston 39 [(80 bar - 104.4 bar) × 2] + [(118.5 bar - 90 bar) × 1] = -48.8 bar + 28.5 bar = -20.3 bar.

[0078] The control piston 39 is pressed towards oil chamber 33 with 20.3 bar times the area unit “1”.

[0079] In a parallel suspension system, if the spring system extends beyond the driving position, the rebound movement is not damped by control pistons 38, 39. This is intentional according to the invention, as precisely this undamped movement, unaffected by the control pistons 38, 39, results in increased ride comfort. This will be explained below.

[0080] A key feature of the damping system 2 is that the damping acts in such a way that it provides a pressure-level-dependent damping effect even at the slightest extension of the spring cylinders. This means that the spring cylinders do not first need to develop an extension speed to generate a damping effect; instead, the speed-independent damper control significantly reduces any increase in speed right from the start. Due to this effective extension damping, the spring system does not oscillate far beyond the driving position. The vehicle weight pushes the spring system back towards the driving position, allowing it to extend slightly beyond this position, as there is no damping influence from the control pistons 38, 39. The subsequent compression movement is triggered solely by the vehicle weight.The increasing spring force counteracts the vehicle weight, and the spring system travels a short distance beyond the driving position in the compression direction. During the return to the driving position, it is damped by the actuating force, now acting again towards oil chambers 67, 68 on the control pistons 38, 39. The respective spring movement around the driving position in a parallel suspension is slowed by the vehicle weight in the rebound direction and by the increasing spring force in the compression direction. Therefore, the damping effect on the spring travel from the compressed position to the driving position is sufficient, as it takes effect immediately upon the start of rebound without delay and immediately counteracts any increase in the rebound speed. The smooth oscillation of the spring movement around the driving position is particularly advantageous, as it prevents acceleration shocks from affecting the vehicle frame and thus also the driver.

[0081] If only the spring cylinder on one side of the vehicle compresses, e.g. spring cylinder 3R, then the following applies: Fig. 20 exemplary pressures shown in damping system 2.

[0082] The force of the adjustment on the control piston 38, 39 is positive when it points towards the oil chambers 67, 68 and negative when it points towards the oil chambers 32, 33. The oil pressures and thus the adjustment forces acting on the control piston 38 are [(213 bar - 118.5 bar) × 2] + [(104.4 bar - 104.4 bar) × 1] = +189 bar.

[0083] The control piston 38 is pressed towards oil chamber 67 with 189 bar times area unit “1”.

[0084] The oil pressures and thus the adjusting forces [(104.4 bar - 104.4 bar) × 2] + [(118.5 bar - 213 bar) × 1] = -94.5 bar act on the control piston 39. The control piston 39 is pushed towards oil chamber 33 with 94.5 bar times area unit “1”.

[0085] When the vehicle suspension on the left side of the vehicle is stationary, the control piston 38 is pushed towards oil chamber 67, thus closing the stepped bore 102 and bore 57. On the right side of the vehicle's suspension, the control piston 39 is pushed towards oil chamber 33.

[0086] In the damping valve 15.1L, the oil flow through bore 57, 102 from oil chamber 67 to oil chamber 30 is closed, and on the right side of the suspension, in the damping valve 15.1R, the oil flow through bore 58, 103 from oil chamber 68 to oil chamber 31 is released. This counter-rotating piston adjustment of the control pistons 38, 39 according to the invention significantly increases suspension comfort.

[0087] Unlike a parallel suspension, where the spring cylinders on both sides of the vehicle compress and rebound in the same direction, in an independent suspension, where only the spring cylinders 3L, 3.1L, 3.2L, 3.3L or the spring cylinders 3R, 3.1R, 3.2R, 3.3R on one side of the vehicle move, the spring movement from the compressed position to the driving position is not damped by the damping control in the damping valve 15.1L, 15.1R. The spring cylinders 3L, 3.1L, 3.2L, 3.3L or the spring cylinders 3R, 3.1R, 3.2R, 3.3R therefore perform a "damping-free" rebound movement. This is particularly advantageous with independent wheel suspension, as the faster rebound movement of the compressed spring cylinder stabilizes the vehicle and prevents the vehicle frame, and therefore the vehicle itself, from tilting sideways.

[0088] If the spring cylinder 3L, 3.1L, 3.2L, 3.3L or the spring cylinder 3R, 3.1R, 3.2R, 3.3R on one side of a vehicle with independent suspension extends beyond the driving position, the rebound movement is increasingly damped by the damping control of the damping valve 15.1L, 15.1R as the extension length increases. The rebound damping in independent suspension is therefore displacement-dependent from the driving position towards the extension. A detailed explanation of the calculation process, as shown above, is omitted, since the direction of movement of the control pistons 38, 39 can be demonstrated using the pressures. Fig. Figure 21 illustrates this for the extension of the spring cylinder 3R, 3.1R, 3.2R, 3.3R of the right spring system.

[0089] How Fig. As shown in Figure 21, the control piston 38 in the damping valve 15.1L moves towards the oil chamber 32 because the oil pressure on surface 78 is 108.5 bar and the oil pressure on the opposite surface 79 is 75 bar. The oil path from oil chamber 67 to oil chamber 30 is therefore not blocked by the closed bores 57 and 102. This control is also very advantageous for driving comfort, as will be explained below.

[0090] When the spring cylinder 3R, 3.1R, 3.2R, 3.3R on the right side of the vehicle, which has extended beyond the driving position, is retracted by the road contour, an increasing spring force acts on the vehicle frame on the right side. This increasing force attempts to induce a lateral roll in the vehicle. Since the damping valve 15.1L does not obstruct the flow from oil chamber 30 to oil chamber 67, as the oil can flow through bores 57 and 102, the damping valve 15.1L does not exert any additional damping force on the spring cylinder 3L, 3.1L, 3.2L, 3.3L on the left side of the vehicle when the suspension compresses from the driving position in the direction of compression. Therefore, the spring system has no "fixed point" on the left side of the vehicle that would act as a "pivot point" for excessive lateral roll.This “yielding pivot point” on one side of the vehicle, together with the controlled rebound damping on the other side of the vehicle, prevents the vehicle suspension system from rocking sideways.

[0091] If the vehicle's suspension system performs an opposing spring movement between the vehicle sides, e.g., the spring cylinder 3L, 3.1L, 3.2L, 3.3L on the left side of the vehicle compresses and the spring cylinder 3R, 3.1R, 3.2R, 3.3R on the opposite right side extends, the following oil pressures will result in the damping system 15.1L, 15.1R, for example, see Fig. 22.

[0092] As already written above, the force of the adjustment on the control piston 38, 39 is positive when it points towards the oil chambers 67, 68 and negative when it points towards the oil chambers 32, 33.

[0093] The oil pressures and thus the adjusting forces act on the control piston 38 [(65 bar - 118.5 bar) × 2] + [(104.4 bar - 188 bar) × 1] = -107 bar + -83.6 bar = -83.6 bar

[0094] The control piston 38 is pressed towards oil chamber 32 with 83.6 bar times area unit “1”.

[0095] The oil pressures and thus the adjusting forces [(188 bar - 104.4 bar) × 2] + [(118.5 bar - 65 bar) × 1] = +167.2 bar + 53.5 bar = 220.7 bar act on the control piston 39.

[0096] The control piston 39 is pressed towards oil chamber 68 with 220.7 bar times area unit “1”.

[0097] Since the flow through bores 57, 58 is strongly influenced by the force of the control pistons 38, 39, there is a significant drop in oil pressure in damping valve 15.1R. This results in a high damping force on the rebound side. In damping valve 15.1L, the flow from oil chamber 30 to oil chamber 67 is not affected. Therefore, the conditions are the same as in an independent suspension, except that on the rebound side, the oil pressure in chamber 31, and thus in spring cylinders 3R, 3.1R, 3.2R, 3.3R, is reduced much more significantly.

[0098] Even in the case of a suspension movement of the spring cylinders, where there is a compression movement on one side of the vehicle and a rebound movement on the other side, the compression side is not influenced by the damping valve and this side therefore offers, as with an independent wheel suspension, no "fixed point" that represents a "pivot point" for an excessive lateral tilting movement.

[0099] As the preceding explanations regarding the damping valves 15L, 15R, 15.1L, 15.1R illustrate, the damping control by the damping valves depends not only on the weight that the spring system must support, but also on the position of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R during spring movement. Both the vehicle's weight and the movement of the spring cylinders alter the oil pressure in the spring system and thus, in damping system 2, the adjusting force on the control pistons 36, 37, 38, 39 and their direction of movement towards the oil chamber 30, 31 or towards the oil chamber 32, 33.

[0100] How Fig. 2 and Fig. Figure 16 shows that the spring cylinder 3L with line 72 is switched via the damping valve 15.1R to the line 69R and thus to the spring cylinder 3R, and the spring cylinder 3R with line 73 is switched via the damping valve 15.1L to the line 69L and thus to the spring cylinder 3L.

[0101] The spring cylinder 3L, 3R has the piston 108, 109 with seal 110, 111 and the piston rod 108.1, 109.1 attached to the piston 108, 109. The seal 110, 111 divides the cylinder chamber into the piston chamber 112L, 112R and the annular chamber 113L, 113R.

[0102] The annular space 113L is hydraulically connected to the piston space 112R without the intermediate damping influence of the damping valve 15.1R and is thus hydraulically synchronized with it, and the annular space 113R is hydraulically synchronized with the piston space 112L without the intermediate damping influence of the damping valve 15.1L.

[0103] As already written and derived above, the load-bearing capacity of a hydraulic cylinder is the oil pressure on the area of ​​the piston 108, 109 minus the oil pressure on the area of ​​the annular space 113L, 113R. In explaining the mode of action of the damping system, it was already stated that the pressure equalization of the piston space 112L, 112R with the annular space 113R, 113L located on the other side of the vehicle achieves an increase in the damping system 2.

[0104] In addition to the load- and cylinder-stroke-dependent damping adjustment, the oil pressure introduced into the annular chamber 113L, 113R also contributes to improved suspension behavior. The pressure clamping of the piston 108, 109, 110, 111 between piston chamber 112L, 112R and annular chamber 113L, 113R dampens any oscillating movement of the piston 108, 110, 109, 111 and thus of the piston rod 108.1, 109.1. Therefore, the suspension behavior improves with a higher clamping force acting on the piston 108, 110, 109, 111.

[0105] The vehicle suspension with damping system, as in Fig. 2 to Fig. As shown in Figure 5, the additional suspension comfort is provided by the weight compensation between the right and left sides of the vehicle. Furthermore, the hydraulic weight compensation circuit provides better pressure clamping of the piston 108, 110, 109, 111 between piston chamber 112L, 112R and annular chamber 113L, 113R, thus more effectively damping any oscillating movement of the piston 108, 110, 109, 111 and consequently of the piston rod 108.1, 109.1, resulting in a significant improvement in the suspension characteristics of the vehicle's suspension.

[0106] As in Fig. As shown in Figures 3 to 5, with just one damping system, 2 several spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R arranged in the vehicle can be damped with the damping valve 15L, 15R, 15.1L, 15.1R in the rebound movement.

[0107] The spring cylinders of a suspension assembly are hydraulically interconnected, and the suspension assembly always contains an even number of spring cylinders, e.g., two, four, six, eight, ten, etc. In the suspension assembly, the piston chambers 112L of the spring cylinders are hydraulically connected to the annular chambers 113R, and vice versa. Fig. 16. Furthermore, the piston chambers 112L, 112R of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R on one side of the vehicle are hydraulically connected, and oil exchange and pressure equalization occur between the piston chambers on that side of the vehicle. The annular chambers 113L, 113R of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R on one side of the vehicle are also hydraulically connected, and oil exchange and pressure equalization occur between the annular chambers on that side of the vehicle. The load-bearing capacity of the spring cylinders in a suspension group is the load-bearing capacity of one spring cylinder multiplied by the number of spring cylinders in the suspension group.

[0108] To minimize flow losses during spring movement of the oil flow between the piston chambers 112L, 112R and the annular chambers 113L, 113R and through the damping device 2 to the pressure accumulator system 1L, 1R, the lines connected to the piston chambers 112L, 112R are connected to the flow distributor 11.1 L, 11.1 R, 13.1L, 13.1R, see Fig. 3 to 5 and Fig. 23 to 25, and according to the invention are combined by flow-optimized V-shaped channels 114 into a channel 115, 116 to which the line 69L, 69R is connected. In the same way, the lines connected to the annular spaces 113L, 113R are connected to the flow distributor 11.2L, 11.2R, 13.2L, 13.2R and, according to the invention, combined by flow-optimized V-shaped channels 114 into a channel 115, 116 to which the line 72, 73 is connected. The pressure storage system 2 is connected to the piping system 4 by flow distributor 12L, 12R, 14L, 14R with V-shaped channels 114 that are combined into a channel 116, 117.

[0109] The oil flow through the V-shaped channels 114 and channels 115, 116, 117 is aerodynamically optimal in both flow directions. To achieve this, the angle β between the adjacent flow channels is preferably a maximum of 15° to 20°. One channel, 115, 116, 117, see Fig. 23 to 25, in the flow distributor, the cross-section is dimensioned so that it is equal to the sum of the cross-sections of the V-shaped channels 114.

[0110] The piston chambers 112L, {112.1 L, 112.2L, 112.3L} and also the piston chambers 112R, {112.1 R, 112.2R, 112.3R} are each hydraulically connected to each other via the flow dividers 11.1L, 11.1R, 13.1L, 13.1R, see Fig. 3 to Fig. 5 and Fig. 16. During spring movement of the spring cylinders, oil equalization occurs between the piston chambers 112L, {112.1L, 112.2L, 112.3L} and 112R, {112.1R, 112.2R, 112.3R}. To enable this equalization of oil volume between the spring cylinders 3L, 3.1L, 3.2L, 3.3L and 3R, 3.1R, 3.2R, 3.3R, respectively, without mechanical stroke limitation of the spring cylinders, the spring travel of the spring cylinders is designed, contrary to other vehicle suspensions, such that the extension travel of the spring cylinders is always greater than the compression travel. Preferably, the total spring travel is dimensioned such that the extension travel is 2 / 3 and the compression travel is 1 / 3 of the total spring travel.Such a suspension travel distribution significantly increases driving comfort, since the freedom of movement of the piston 108, 109 in the spring cylinder is so great that when driving over an obstacle, a hill or a hole, the piston 108, 109 does not hit the end of the cylinder tube during a rebound movement, which would cause a metallic shock to the vehicle frame.

[0111] In an advantageous embodiment, the total spring travel of a spring cylinder 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R is between 500 mm and 800 mm.

[0112] How Fig. As shown in Figure 1, the spring cylinders 3L and 3R are connected to the reservoir system 1L and 1R via the line system 4 and the damping system 2. During spring movement, the spring cylinders 3L and 3R push oil into the reservoir system 1L and 1R, or the reservoir system pushes oil back into the spring cylinders. As mentioned above, the spring cylinders have very large diameters and strokes. These large diameters are necessary to support the vehicle's load when fully loaded. Therefore, during the stroke of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, and 3.3R, a large volume of oil is forced into the reservoir system 1L and 1R. Consequently, the reservoir system 1L and 1R must be dimensioned accordingly. Fig. 3 to Fig. Figure 5 shows that several spring cylinders (3L, 3.1L, 3.2L, 3.3L, or 3R, 3.1R, 3.2R, 3.3R) on one side of the vehicle act on the storage system (1L or 1R), which necessitates an even larger storage system (1L, 1R). This additional enlargement results from the fact that the spring system has to bear less load when the vehicle is unloaded than when it is loaded, as explained below.

[0113] The spring force of the spring system is generated by nitrogen pressure in the storage system 1L, 1R. When the nitrogen is compressed, the nitrogen pressure increases, and the spring system has a higher load-bearing capacity. When the nitrogen expands, the nitrogen pressure decreases, and the spring system has a lower load-bearing capacity. The load-bearing capacity therefore changes with the level of nitrogen pressure and thus with the volume of nitrogen.

[0114] The storage system 1L, 1R consists of a gas chamber 118 filled with nitrogen and an oil chamber 121 separated by a piston 119 with seal 120, to which the piping system 4 is connected, see Fig. 26. A piston rod 122 is attached to the piston 119, extending through the oil chamber 121 into a second gas chamber 123, which is also filled with nitrogen. The gas chamber 123 is sealed against the oil chamber 121 by a seal 124.

[0115] The size of the gas chamber 118 with a nitrogen pressure P1.1 is determined by multiplying the size under load by the ratio of loaded to empty spring load. This loaded to empty ratio can easily be on the order of ten or more.

[0116] The expansion of gas space 118 is counteracted by a nitrogen pressure P1.2 in gas space 123, since the piston rod 122 is pushed into gas space 123, thereby reducing the nitrogen volume in gas space 123 and thus increasing the nitrogen pressure P1.2, thereby increasing the pushing force of the piston rod 122 towards gas space 118. For gas space 118, this means that when the spring load changes from loaded to empty, it does not expand as much, because with less spring load the piston rod 122 is pushed further into gas space 123, thereby increasing the counterweight on the piston 119 and thus on gas space 118.

[0117] The nitrogen pressure P1.1 in the gas chamber 118 must therefore counteract the load from the oil pressure in the oil chamber 121 plus the load from the sliding force of the piston rod 122. Thus, if the oil pressure in the oil chamber 121 decreases and the nitrogen pressure P1.1 in the gas chamber 123 increases, the nitrogen pressure P1.1 in the gas chamber 118 changes less when the vehicle's load changes. The expansion of the nitrogen is less, and therefore the increase in the size of the gas chamber 118 is also smaller. This advantageous design of the storage system 1L, 1R allows the dimensions of the storage system to be drastically reduced, which has a positive effect on the available space in the vehicle. For example, with a large weight change from loaded to empty, e.g., a ratio of 10:1, the counter-pressure of chamber 123 halves the required nitrogen volume in the gas chamber 118.

[0118] Despite the counterforce from the nitrogen pressure P1.2 in space 123, the length of the storage system 1L, 1R can be approximately seven meters or more in vehicles with a total weight of up to 600 tons, which makes installation in a vehicle very difficult and also makes the production of such long storage systems extremely complex and expensive.

[0119] Despite significantly higher manufacturing costs, the storage system 1L, 1R according to the invention is designed such that the storage system 1L, 1R consists of several individual storage units 1.1L, 1.2L, 1.3L, 1.4L, 1.1R, 1.2R, 1.3R, 1.4R, see Fig. 4 and Fig. 5.

[0120] A key feature of the storage system 1L, 1R is that the individual storage units 1.1L, 1.2L, 1.3L, 1.4L from storage system 1L and the individual storage units 1.1R, 1.2R, 1.3R, 1.4R from storage system 1R each function like one large individual storage unit 1L and 1R across the entire loading area of ​​the vehicle and also during the spring movements of the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R, and do not use the individual storage units sequentially.

[0121] The individual gas pressure accumulators 1.1L, 1.2L, 1.3L, 1.4L 1.1R, 1.2R, 1.3R, 1.4R of the storage system 1L, 1R are used sequentially because each pressure accumulator is filled with a different filling pressure p1, p2, p3, p4, p5, p6, see Fig. 27, thus a suspension system is created in which the spring force increases exponentially due to the nitrogen compression in the first pressure accumulator up to a maximum force and then the exponential nitrogen compression in the next pressure accumulator connects to the spring force of the previous pressure accumulator, see Fig. 28. The force curve of the vehicle suspension exhibits a step-like increase in force during the compression movement of the spring cylinders and a similar step-like decrease in force during the rebound movement of the spring cylinders. Such a suspension behavior of the vehicle's suspension system leads to vehicle handling that is no longer easily controllable by the driver, resulting in an increased risk of accidents.

[0122] For safe handling and comfortable driving, it is advantageous if there is no step-like force curve in the vehicle suspension during suspension movement. Advantageously, the vehicle suspension should provide a uniform spring action over a wide range of vehicle loads. This is achieved by filling the gas chamber 118 of each individual accumulator (1.1L, 1.2L, 1.3L, 1.4L and 1.1R, 1.2R, 1.3R, 1.4R) with the same filling pressure P1.1 via a gas filling valve 124, corresponding to the filling pressure P1.1 of a single large accumulator (1L, 1R). In contrast, the gas space 123 of each individual storage tank 1.1L, 1.2L, 1.3L, 1.4L and 1.1R, 1.2R, 1.3R, 1.4R is filled via a gas filling valve 125 with a filling pressure P7, which is different from the filling pressure P1.2 of a large gas space 123 of an individual storage tank 1L, 1R, see Fig. 26 and Fig. 29. This is necessary so that the shorter stroke of the piston 119 of each individual storage unit 1.1L, 1.2L, 1.3L, 1.4L and 1.1R, 1.2R, 1.3R, 1.4R has the same pressure change in the gas space 123 as a longer individual storage unit 1L, 1R. Without this pressure adjustment of the filling pressure P7 of the gas space 123, the storage system 1L, 1R does not operate correctly, since a filling pressure equal to that in a large gas space 123 results in a different nitrogen compression in the gas space 123 due to the different stroke of the piston rod 122.

[0123] When filling gas chamber 118 and gas chamber 123 with nitrogen, it must be taken into account that an unavoidable tolerance in the nitrogen filling leads to a difference in the pressure increase in gas chamber 118 and gas chamber 123, and as a result, the gas accumulators do not operate synchronously, which can lead to a disturbance in the force curve of the vehicle suspension.

[0124] In an improved embodiment of the storage system 1L, 1R, designed for adjusting the filling pressure and for subsequent servicing, the gas chamber 118 and the gas chamber 123 are interconnected in the storage system 1L, 1R with individual storage units 1.1L, 1.2L, 1.3L, 1.4L and 1.1R, 1.2R, 1.3R, 1.4R, according to the invention. Fig. Figure 30 shows this improved storage system 1L, 1R as an example with three individual storage units.

[0125] The gas spaces 118 are connected to each other via line 126 and the nitrogen is filled into the gas spaces 118 together via gas filling valve 127, and the gas spaces 123 are connected to each other via line 128 and the nitrogen is filled into the gas spaces 123 together via gas filling valve 129.

[0126] Since gas spaces 118 and 123 are connected by lines 126 and 128 respectively, it is ensured that the same nitrogen set pressure is present in all gas spaces 118 and that the set pressure is also the same in all gas spaces 123, and that any unavoidable filling tolerance does not lead to different compression in gas spaces 118 and 123 of the individual storage tanks 1.1L, 1.2L, 1.3L, 1.4L and 1.1R, 1.2R, 1.3R, 1.4R of the storage system 1L, 1R.

[0127] Due to the parallel connection of all gas spaces 118, all spaces together act like the one large gas space 118 of a single, extremely large storage tank 1L, 1R, and the same is true for all gas spaces 123. The filling pressure P1.1 of the gas spaces 118 is equal to the filling pressure of an extremely large gas space 118 of a large storage tank 1L, 1R, and the same applies to the parallel-connected gas spaces 123 with the filling pressure P1.2.

[0128] The driving behavior of a vehicle is not only influenced by the actual suspension system with the shock absorbers, but also by the vehicle wheels with the tire suspension, which have a noticeable influence on driving comfort.

[0129] As mentioned above, tires with a large air volume, as required for tires with very high load capacities, have the disadvantage that they provide inherent suspension that cannot be damped. With such tires, the vehicle tends to exhibit oscillations.

[0130] Based on the understanding that the vehicle suspension system described above, with its "damping-free" compression, transmits extremely few, or barely perceptible, shocks from the road surface to the vehicle frame, the vehicle wheel 6L, 6R is designed in such a way that it does not cause or stimulate any lifting oscillation and yet allows a higher load. Fig. 31 and Fig. Figure 32 shows the vehicle wheel 6L, 6R according to the invention for vehicles with a payload of 100 tons and more.

[0131] The vehicle wheel 6L, 6R is designed such that it consists of a disc 130 and a ring 131 firmly connected to it with a vulcanized or bonded wear-resistant coating 132, see Fig. 1 to 5 and Fig. 31. The disc 130 has a central hole 133 and bores 135 on a radius 134 for attaching the vehicle wheel to the vehicle axle. The coating 132 is the running surface of the wheel 6L, 6R, and in a preferred embodiment, the coating 130 has a layer thickness of approximately 100 mm and consists of an abrasion-resistant rubber compound.

[0132] In an improved version, the vehicle wheel 6L, 6R is provided on the tread surface with a coating 136 which consists of at least two superimposed and firmly bonded layers 137, 138, see Fig. 32. The inner coating 137, vulcanized or bonded to the ring 131, is softer than the outer, more wear-resistant running layer 138. In a preferred embodiment, the inner, softer coating 137 is approximately 40 mm thick, and the outer, harder, wear-resistant coating is 60 mm thick. Because the outer coating 138 can be pressed slightly into the softer coating 137, the wear of the outer coating 138 is reduced. This reduction in wear occurs because, when driving over, for example, a stone, the outer layer 138 presents practically no rigid surface to the stone, but yields slightly, which drastically reduces the impact on the running surface and thus the force causing wear.

[0133] For changing a wheel, it would be advantageous if an axle could be lifted using the hydraulic spring system as described above. However, since such lifting of an axle requires shut-off devices to be connected between the pressure accumulator system 1L, 1R and the spring cylinders 3L to 3R in the line system 4, and these shut-off devices result in very high flow losses during spring movement, this solution is not implemented in the spring system according to the invention due to the drastic reduction in ride comfort. Instead, a different lifting device for the axles is installed in the vehicle in conjunction with the spring system described above, as shown in [reference to figure]. Fig. 33 and Fig. 34 shown.

[0134] For each vehicle wheel 6L, 6.1L, 6.2L, 6.3L, 6R, 6.1R, 6.2R, 6.3R, hooks 140 are rotatably mounted in a bearing 139 in the vehicle frame 7, and these are held in their folded position by bolts slidably mounted in the vehicle frame, hereinafter referred to as sliding bolts 141, see Fig. 33. The sliding bolt engages either in a hole in the hook 140 or under a pin 142, as shown in Fig. Figure 33 shows that a bracket 144 with bolt 145 is attached to axle 143. Together with the bracket 144 and the hook 140 which engages the bolt 145, the axle can be held in a raised position so that a wheel 6L, 6.1L, 6.2L, 6.3L, 6R, 6.1R, 6.2R, 6.3R does not touch the road surface even when the vehicle's suspension compresses. Fig. 34.

[0135] To connect the axle 143 to be lifted to the hook 140, the vehicle is lowered to its maximum position. For this purpose, oil from the suspension system is drained via line 151 into an oil tank 152 through a hydraulic block 146 with electric switching valves 147, 148, and the vehicle lowers to its lowest possible position (see figure). Fig. 35. The sliding bolt 141 is moved so that the hook 140 can move downwards and engage the bolt 145. With the aid of the oil pump 149, the vehicle is raised again after the hook 140 has engaged in the holder 144 until it is at driving level.

[0136] To raise the vehicle back into the driving position, oil is pumped into the hydraulic block 146 via line 150 using the oil pump 149, and by electrically switching the switching valves 147, 148 the oil flow to line 153, 154, connected to line 69L, 69R, is released, and the oil flows into the line system 4, the spring cylinders 3.1L, 3.2L or 3.1R, 3.1R and into the storage system 1L or 1R.

[0137] The in Fig. 1 to Fig. The spring system shown in Figure 5 is for one axle or for several interconnected axles. A wheeled vehicle requires at least two independent axles, i.e., a front axle and a rear axle. Both the front and rear axles can consist of a single axle or several interconnected axles. To raise and lower a vehicle, oil is pumped into the front axle, the rear axle, and also the right and left sides of the vehicle independently of each other using oil pump 149, or drained from these axles into tank 152.

[0138] To ensure that the spring system remains absolutely leak-proof during travel, the oil pump 149 does not operate directly via the switching valves 147, 148 on line 153, 154, but rather a hydraulically unlockable check valve 155, 156 is installed between switching valve 147, 148 and line 153, 154 in line system 4, see Fig.35. With a releasable check valve, oil can flow through in one direction, while in the other direction, the valve must be opened by hydraulic pressure. When oil is pumped into the spring system, it flows into line 153, 154 without the check valve 155, 156 being opened. When oil is drained from the spring system, the pump pressure of the oil pump 149 opens the check valve 155 or check valve 156 via a control line 157 or control line 158 through the switching valve 147 or switching valve 148, and the oil can then flow through line 153 or line 154, the check valve 155 or check valve 156, the switching valve 147 or switching valve 148, and line 151 into the oil tank 152.

[0139] Sensors (not shown) on the vehicle determine whether the desired ride height is reached at all axles during raising or lowering. Depending on the sensor signal, valves 147 and 148 are switched so that oil either flows into line 153 or line 154, or is diverted through line 153 or line 154 via control block 146 and line 151 into oil tank 152. This allows the vehicle to be raised and lowered independently on each side, as well as at the front and rear.

[0140] Raising or lowering the vehicle to its driving level is necessary, for example, after a load change. With a lighter load, the nitrogen in the storage system 1L, 1R expands, forcing oil from the storage system into the spring cylinders, thus raising the vehicle. Conversely, with a heavier vehicle, oil is forced from the spring cylinders into the storage system 1L, 1R, causing the vehicle to lower. When the vehicle is returned to its driving level after a load change, this is accomplished using the pump pressure of the oil pump 149. This pump pressure also opens the check valves 17 and 18 in the valve unit 16, and the changed system pressure of the spring system after a load change is stored in pressure accumulators 19, 20.The control of the damping system 2 automatically adjusts to the different vehicle weight in the event of a change in oil pressure after a lifting or lowering operation, and thus has a rebound damping controlled according to this weight and the type of suspension in the case of independent wheel, parallel or counter-rotating suspension.

[0141] Unlike a steel or air suspension, where the dampers are connected in parallel to the vehicle's suspension, in the spring system described above, the damping is connected in series with the suspension element, storage system 1L, 1R. This series connection of the suspension and damping arises because the suspension is provided by the nitrogen volume in storage system 1L, 1R, and the pressure of the nitrogen, combined with oil, is transmitted through damping system 2 to the spring cylinders 3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R. List of reference symbols 1L, 1.1L, 1.2L, 1.3L,1.4L, 1.5L pressure storage system 1R, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R Pressure storage system 2 Damping system 3L, 3.1L, 3.2L, 3.3L, spring cylinder 3R, 3.1R, 3.2R, 3.3R spring cylinder 4. Pipe system 5 elastic coating 6L, 6.1L, 6.2L, 6.3L, vehicle wheel 6R, 6.1R, 6.2R, 6.3R Vehicle wheel 7 Vehicle frames 8L, 8.1L, 8.2L, 8.3L, Lager 8R, 8.1R, 8.2R, 8.3R bearings 9L, 9.1L, 9.2L, 9.3L, swing 9R, 9.1R, 9.2R, 9.3R Swingarms 10 Direction of travel 11.1L, 11.2L, 11.1R, 11.2R Flow distributor 12L, 12R flow distributor 13.1L, 13.2L, 13.1R, 13.2R Flow distributor 14L, 14R Flow distributor 15L, 15.1L, 15R 15.1R Damping valve 16 Valve unit 17, 18 Double check valve 19, 20 Pressure accumulator 21, 22, 23, 24, 25, 26 Hydraulic line 27, 28L, 28R, 29 Hydraulic line 30, 31, 32, 33 Oil room 34, 35 Hydraulic line 36, 37, 38, 39 Control piston 40, 41 Surface area 43, 44 Seal 45, 46 mushroom-shaped piston part 47, 48 Floor 49, 50 pot 51, 52 socket 53, 54 Hydraulic fitting 55, 56 shoulder 57, 58, 59, 60 bore 61, 62 disc 63, 64 basket 65, 66 Hydraulic fitting 67, 68 Oil room 69L, 69R, 70, 71, 72, 73 Hydraulic line 74, 75 Hydraulic fitting 76, 77 Circumference increase 78, 79, 80, 81 Piston area 82, 83, 84, 85, 86, 87 Seal 88, 89, 90, 91 Oil room 92, 93, piston area 94, 95 mushroom-shaped piston part 96, 97 piston area 98, 99 ring area 100, 101 Piston part conical 102, 103 Step-shaped bore 104,105 bore Levels 106 and 107 108, 109 Piston Spring Cylinder 108.1, 109.1 Piston rod spring cylinder 110, 111 Seal 112L, 112.1L, 112.2L, 112.3L piston space

Claims

[1] Vehicle suspension system for multi-axle wheeled vehicles, with a hydraulic line system (4) and a damping system (2) integrated into this line system (4), as well as a pressure accumulator system (1L, 1R, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) connected to this damping system (2) by a line system (4) and spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R) also connected to the line system (4), wherein the damping system (2) has damping valves (15L, 15R), in which an oil chamber (67, 68) is connected to the pressure accumulator system (1L, 1R, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) and an oil chamber (30, 31) is connected to the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R), characterized bythat the damping valve (15L, 15R) has a control piston (36, 37) and the control piston (36, 37) has a mushroom-shaped piston part (45, 46) which can close or open a bore (57, 58) in a pot (49, 50) for oil flow and furthermore in the damping valve (15L, 15R) there is a movable disc (61, 62) which can close or open a second bore (59, 60) in the pot (49, 50) for oil flow and that the mushroom-shaped piston part (45, 46) is arranged in relation to the bore (57, 58) in such a way that when oil flows through a hydraulic screw connection (65, 66) in the damping valve (15L, 15R) to the oil chamber (30, 31), it closes the bore (57, 58) closes and when oil flows through a hydraulic screw connection (53, 54) in the damping valve (15L, 15R) into the oil chamber (67, 68), this oil flow can lift the mushroom-shaped piston part (45, 46) so that oil flows through the bore (57, 58) from the oil chamber (67, 68) into the oil chamber (30,31) is possible and the disc (61, 62) enables the oil flow from the oil chamber (30, 31) to the oil chamber (67, 68) through the bore (59, 60) by lifting the disc (61, 62) when oil flows through the hydraulic screw connection (65, 66) to the oil chamber (30, 31) and when oil flows through the hydraulic screw connection (53, 54) into the oil chamber (67, 68), the disc (61, 62) blocks the oil flow through the bore (59, 60), wherein the disc (61, 62) and also the mushroom-shaped piston part (45, 46) are moved by the flow direction of the oil flow, but in contrast to the disc (61, 62), the mushroom-shaped piston part (45, 46) is additionally controlled in its stroke movement by the control piston (36, 37) and thus a controlled oil throttling is present during the oil flow from the oil chamber (67, 68) to the oil chamber (30, 31) and no desired oil throttling is generated during the oil flow from the oil chamber (30, 31) to the oil chamber (67, 68) through the disc (61, 62),whereby this different flow influence by the mushroom-shaped piston part (45, 46) and the disc (61, 62) creates a damping system which, upon compression of the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R), i.e., a deflection, does not produce any desired oil throttling, i.e., no damping, due to the oil flow from the oil chamber (30, 31) to the oil chamber (67, 68), but upon expansion of the spring cylinders (3L, 3.1L, 3.2L, 3.3L, 3R, 3.1R, 3.2R, 3.3R), i.e., during rebound, a controlled oil throttling, i.e., a controlled damping, is present due to the oil flow from the oil chamber (67, 68) to the oil chamber (30, 31). is., [2] Spring system according to claim 1, characterized bythat the damping system (2) with damping valve (15L, 15.1L, 15R, 15.1R) in the case of independent wheel suspension is controlled by the static oil pressure stored in the pressure accumulator (19, 20) and oil chamber (32, 33) depending on the vehicle load and this static oil pressure is guided into the oil chamber (32, 33) by means of a hydraulic line (21, 23) and acts on the surface (92, 93) of the control piston (38, 39) and that the oil pressure in the oil chamber (30, 31) acts on the annular surface (98, 99) of the control piston (38, 39). [3] Spring system according to claim 1 or 2, characterized by that opposite the control piston (38, 39) there is a mushroom-shaped piston (94, 95) which is displaceable relative to the control piston (38, 39) and the oil pressure from chamber (30, 31) acts on this displaceable mushroom-shaped piston (94, 95) and, on the other hand, the oil pressure from chamber (67, 68) acts on it through the bore (57, 58). [4] Spring system according to one of claims 1 to 3, characterized bythat the oil pressure in oil chamber (30) is guided via hydraulic line (27) into oil chamber (91) and this oil pressure acts on surface (81) of control piston (39) and the oil pressure in oil chamber (31) is guided via hydraulic line (29) into oil chamber (90) and acts on surface (79) of piston (38) and the oil pressure in oil chamber (32) is guided via hydraulic line (28R) into oil chamber (89) and there acts on surface (80) of control piston (39) and the oil pressure from oil chamber (33) is guided via hydraulic line (28L) into oil chamber (88) and there acts on surface (78) of control piston (38). [5] Vehicle suspension system according to one of claims 1 to 4, characterized bythat on the control piston (38, 39) at one end there is a surface (92, 93) and opposite this surface (92, 93) there is a conical piston part (100, 101) and this conical piston part (100, 101) projects into a stepped bore (102, 103) of the mushroom-shaped piston part (94, 95) and the conical piston part (100, 101) when it rests against a step (106, 107) in the stepped bore (102, 103) can close it so that no oil flows through the stepped bore (102, 103) and when the conical piston part (100, 101) does not rest against the step (106, 107) a small oil flow through the stepped bore (102, 103) and this oil flow can flow out of the stepped bore (102, 103) through a bore (104, 105) in the mushroom-shaped piston part (94, 95) and controlled by this inventive design of the control piston (38, 39) and the mushroom-shaped piston part (94, 95), a small oil flow through the bore (57,58) from the oil chamber (67, 68) to the oil chamber (30, 31) and this slight oil flow between the oil chamber (30, 31) and the oil chamber (67, 68) ensures that the static driving position is reliably reached and maintained after a spring movement, which is a significant increase in driving comfort, since without this slight oil flow an oil pressure difference in the static driving position between the oil chamber (30, 31) and the oil chamber (67, 68) can arise, which would lead to a slight tilt of the vehicle. [6] Vehicle suspension system according to one of claims 1 to 5, characterized bythat elastic seals (43, 44, 82, 83, 84, 85, 86, 87) are arranged on the outer surface of the control piston (36, 37, 38, 39) in the damping valve (15L, 15.1L, 15R 15.1R), which are prestressed in such a way that the seals (43, 44, 82, 83, 84, 85, 86, 87) on the control piston (36, 37, 38, 39) have a restoring force towards the oil chamber (32, 33) and by means of this restoring force on the control piston (36, 37, 38, 39) an opening of the bore (57, 58, 102, 103, 104, 105) can be achieved and thus between the spring cylinder (3L, 3.1L, 3.2L,_3.3L, 3R, 3.1R, 3.2R, 3.3R) and the oil chamber (67, 68) assigned to the pressure storage system (1L, 1R), a small oil flow is possible between the oil chamber (30, 31) and the oil chamber (67, 68). [7] Vehicle suspension system according to one of claims 1 to 6, characterized bythat the stroke of the control piston (36, 37, 38, 39) in the damping valve (15L, 15.1L, 15R 15.1R) for adjusting the oil throttling with an oil flow direction from the oil chamber (67, 68) to the oil chamber (30, 31) in the damping valve (15L, 15.1L, 15R 15.1R) is between 0.5 mm and 1.5 mm and in a preferred embodiment is 1.0 mm. [8] Vehicle suspension system according to one of claims 2 to 7, characterized by that the oil chamber (32, 33) assigned to the control piston (36, 37, 38, 39) is acted upon by the static oil pressure in the pressure accumulator (19, 20) and a line (21, 23) between the pressure accumulator (19, 20) and the oil chamber (32, 33) enables the control piston (36, 37, 38, 39) to move freely in order to throttle the oil flow through the damping valve (15L, 15.1L, 15R 15.1R) and is not prevented from making a stroke movement by a chambered oil volume in the oil chamber (32, 33). [9] Vehicle suspension system according to one or more of claims 1 to 8, characterized bythat the pressure accumulator system (1L, 1R) consists of at least one pressure accumulator (1L, 1R) and, in a preferred embodiment, of several pressure accumulators (1.1L, 1.2L, 1.3L,1.4L, 1.5 L, 1.1R, 1.2R, 1.3R,1.4R, 1.5R). [10] Vehicle suspension system according to one or more of claims 1 to 9 characterized bythat each pressure accumulator (1L, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1R, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) of the pressure accumulator system (1L, 1R) has a piston (119) with a piston rod (122), wherein the pressure accumulator (1L, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1R, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) has nitrogen-filled gas spaces (118, 123), between which the piston (119) with a piston rod (122) is arranged, wherein the piston rod (122) projects into the gas space (123) and thus the end face of the piston rod (122) is pressurised with the nitrogen pressure of the gas chamber (123) and the gas chamber (118) pressurises the piston side of the piston (119) opposite the piston rod (122) with nitrogen pressure, and the respective gas chamber (118) of the pressure accumulators (1L, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1R, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R)_are connected to one another by line (126) and the respective gas chamber (123) of the pressure accumulators (1L, 1.1L, 1.2L, 1.3L, 1.4L, 1.5L, 1R, 1.1R, 1.2R, 1.3R, 1.4R, 1.5R) are connected by lines (128), and the gas space (118) can be filled via filling valves (127) and the gas space (123) can be filled via filling valve (129). [11] Vehicle suspension system according to one of claims 1 to 10, characterized by that in order to compensate for a different weight of a right and left vehicle side, a hydraulic pressure from a piston chamber (112L, 112.1L, 112.2L, 112.3L) of the spring cylinder (3L, 3.1L, 3.2L, 3.3L) is passed via a line (69L, 73) into the annular chamber (113R, 113.1R, 113.2R, 113.3R) of the spring cylinder (3R, 3.1R, 3.2R, 3.3R) and the hydraulic pressure from a piston chamber (112R 112.1R, 112.2R, 112.3R) via a line (69R, 72) into the annular chamber (113L 113.1L, 113.2L, 113.3L) of the spring cylinder (3L, 3.1L, 3.2L, 3.3L). [12] Vehicle suspension system according to one of claims 1 to 11, characterized bythat flow distributors (11.1L, 11.2L, 11.1R, 11.2R, 12L, 12R, 13.1R, 13.2R, 13.1L, 13.2L, 14L, 14R) with a flow channel (114, 115, 116, 117) are installed in the hydraulic line system (4) to minimize the flow resistance. [13] Vehicle suspension system according to claim 12, characterized by that the flow channel (114) is V-shaped and has an angle (β, β / 2) to each other which lies between 15° and 20°.

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