EP2179176B1 - Hydraulic system with an adjustable hydrostatic machine - Google Patents

Description

The invention relates to a hydraulic system with an adjustable hydrostatic machine, a co-operating with the hydrostatic machine actuator and a control valve, wherein the actuating device has at least a first actuating piston, which is acted upon on a control piston surface with a control pressure, and wherein the control valve has a control element, which is acted upon in the direction of a first end position of the valve with a dependent on the position of the first actuating piston force and which is acted upon for adjusting a control valve position at a second end with a force.

From the DE 199 49 169 C2 a hydrostatic system with an arranged in an open circuit variable displacement hydraulic pump according to the preamble of claim 1 is known. The adjusting mechanism of the hydraulic pump is actuated by an adjusting device. The adjusting device has an actuating piston which defines a control pressure chamber formed in a control cylinder. The control pressure prevailing in the signal pressure chamber is regulated by a control valve. The control pressure chamber is variably connectable with a delivery-side working line of the hydraulic pump or a tank volume for setting a control pressure acting on the control piston surface of the control piston via the control valve.

The control valve has a valve piston as a valve piston which is acted upon at a first end with the force of a feedback spring. The feedback spring supported on the one hand at one end of the valve piston and on the other hand on the acted upon by the control pressure piston surface of the actuating piston. The force generated by the feedback spring on the valve piston is thus dependent on the position of the actuating piston. In the opposite direction acts on the valve piston, the force of an electromagnet. As a function of a control signal for the electromagnet, a proportional adjustment of the actuating piston is thus achieved as a function of a control signal of the proportional magnet.

Although the proposed hydrostatic system allows due to the arrangement of the feedback spring on the one hand and the Proportiönalmagneten other hand, a simple integration into a pump unit. However, it is disadvantageous that a center position of the actuating piston is not possible with vanishing control signal and a variable in two opposite directions hydrostatic machine can only be brought by generating a control force by the proportional solenoid in its neutral position.

The EP 0 849 468 A2 shows an adjusting device for adjusting the delivery volume of an axial piston pump. The adjusting device comprises an actuating piston adjoining the delivery volume of the axial piston pump in a force-locking manner. Further, a control valve is provided, which has a valve piston on which a connecting member bears non-positively. On the connecting member engages a return spring, which is supported on a first spring plate. Between the adjusting piston and the non-positively connected to the valve piston of the control valve connecting member, a feedback spring is arranged, which acts on a voltage applied to the connecting member second spring plate. Further, a Ausschwenkfeder is provided, which is clamped between the first spring plate and a substantially formed as a hollow cylinder stop element.

The EP 0 859 380 A2 shows an electromagnetic Doppelhubmagneten.

The WO 03/014570 A1 shows a valve block for a control device, in particular for a hydrostatic machine, with a pressure control valve and a flow control valve, each having a valve piston and a return spring.

It is therefore an object of the invention to provide a hydrostatic system which, starting from a rest position with vanishing control signal in two opposite directions is pivotable.

The object is achieved by the hydraulic system according to the invention with the features of claim 1.

The hydraulic system according to the invention has an adjustable hydrostatic machine, an adjusting device interacting with an adjusting mechanism of the hydrostatic machine and a control valve.

The adjusting device has at least one actuating piston, which is acted on at least on an actuating piston surface with a control pressure. The control valve has as a control element to a valve piston which is acted upon by a dependent of the position of the actuating piston force. To set the desired position of the adjusting mechanism, the control element can be acted upon by a control force which can be generated by actuating means.

By providing an acting on the control element and counter to the dependent of the position of the actuating piston force acting counterforce even with vanishing control signals, when no control force is generated on the control element, a balance of forces on the control valve in a rest position of the first actuating piston can be achieved. Since the control force generated by the actuating means is directed in the same direction with the counterforce or contrary to her, can be deflected in two opposite directions by the actuating means, starting from the determined by the reaction force rest position of the actuator with vanishing control signal. The hydrostatic system according to the invention therefore has the advantage that not only an adjustment of a hydraulic pump between, for example, a zero delivery volume and its maximum delivery volume is possible, but that also in opposite directions pivotable hydrostatic machines, such as provided for two flow directions pumps or a pump / Motor unit, as used in braking energy recovery devices can be controlled in a simple manner.

In the dependent claims advantageous developments of the hydraulic system according to the invention are carried out.

The counterforce is preferably generated by a spring and the control force is preferably continuously adjustable in both directions of force.

The actuating means preferably comprises two, acting in opposite directions on the valve piston proportional solenoids. Such proportional magnets are simple and inexpensive available.

In particular, it is advantageous that a simple integration of the control valve in a hydrostatic machine is possible by controlling both actuating directions by only one actuator as actuating means, in which the acted upon by the position-dependent force side of the control element is not accessible.

The actuating means is preferably designed according to an alternative embodiment as Doppelhubmagnet with a first winding and a second winding. In this case, a force is generated by each of the windings in each case one of the opposite directions of force. With such an actuator, by applying the first or the second winding, the direction of force and also the magnitude of the force can be predetermined in a simple manner by two individual control signals. In particular, it is possible to control the hydrostatic system in a simple manner using an electronic control unit. The use of only one actuator for both directions of force also has the advantage that the line and plug expenses can be kept small.

A particularly simple construction is achieved if the feedback spring is arranged between the actuating piston and the valve piston.

Furthermore, it is advantageous that the hydrostatic machine is set with vanishing control force and an equilibrium of forces between the opposing force and the dependent of the position of the actuating piston force on the control element to a non-zero delivery volume. Such an adjustment in the rest position of the hydrostatic system to a delivery volume other than zero ensures that after commissioning and when the control signal disappears, a minimum delivery volume is already conveyed through the hydrostatic machine. Because of this minimum delivery volume, there is already a slight increase in pressure in the hydraulic system which can be used to actuate the hydrostatic machine actuator. By the said setting in the rest position is further avoided that occur in the hydrostatic machine intake problems due to wrong pivot side.

The adjusting device preferably has, in addition to the actuating piston, which acts on the hydrostatic machine from a rest position in the direction of a first end position with a restoring force, another, also referred to opposing piston piston, the hydrostatic machine in the opposite direction, starting from the rest position in the direction a second end position acted upon by a force. Such Distribution of the adjusting device in a first actuating piston for a first direction of movement and a second actuating piston for a second direction of movement has the advantage that the available space can be used particularly efficiently, for example in a hydrostatic axial piston machine. Thus, it is possible to arrange the two adjusting pistons and the adjusting cylinders, in which the adjusting pistons are arranged, on opposite sides of a pivoting cradle or swash plate. By the respective actuating piston only compressive forces are to be transmitted to the pivoting cradle, which simplifies the mechanical coupling between the actuating piston and the hydrostatic machine.

The limited by the piston surface of the actuating piston control pressure chamber is preferably connected via the control valve with a control pressure source. In this case, the control pressure source is connected in particular via a connecting line in which a pressure reducing valve is arranged with the control valve. The connecting line with the pressure reducing valve provided therein has the advantage that by means of the pressure reducing valve, a reduced compared to the working pressure input pressure is generated. This can be used as a control pressure source, for example, depending on different operating situations of the hydrostatic system pressure. In particular, the delivery pressure of the hydrostatic machine may be provided as a control pressure source. Furthermore, it is particularly advantageous when using the pressure reducing valve in the connecting line to realize the adjusting device by a first adjusting piston and a second adjusting piston. In this case, the second actuating piston has a smaller compared to the first actuating piston piston surface, the immediate is acted upon by the working pressure of the hydrostatic machine. By reducing the control pressure obtained from the working pressure of the hydrostatic machine with simultaneously larger piston area of the first actuating piston is achieved that the adjusting mechanism of the hydrostatic machine is hydraulically clamped at any time. Nevertheless, the maximum possible actuating pressures of the first actuating piston are reduced compared to the actual working pressure. In particular, thereby the actuating means, but also the control valve is only exposed to a lower pressure load.

The control pressure source is preferably a working line of the hydrostatic machine or a further pressure medium source. The working line or the pressure medium source are connected via a shuttle valve to the connecting line. The possibility of alternatively providing a further pressure medium source in addition to the working line as an actuating pressure source has the advantage that even when the working line pressure disappears, a control pressure sufficient for actuating the hydrostatic system is available. This is particularly advantageous when the hydrostatic system is used in the recovery of kinetic energy by storing the energy in a hydrostatic accumulator. During operation, it may happen that the hydrostatic accumulator is completely emptied. In this case, a sufficient actuating pressure in the first control pressure chamber can be generated by the further pressure medium source, through which the hydrostatic machine can be set to a higher delivery volume in a subsequent storage of energy.

The control valve is preferably designed as a 3/3-way valve. The use of a 3/3-way valve makes it possible to additionally provide a neutral position of the control valve between the two end positions. This neutral position is determined by a defined control valve position. In this neutral position of the control valve, the actuating pressure chamber, which is limited by the actuating piston surface of the first actuating piston, is preferably throttled with a tank volume and throttled to the control pressure source.

In the first end position of the control valve, the control pressure chamber is connected to the control pressure source. In the opposite second end position of the control valve, however, the limited by the control piston surface of the first control piston actuating pressure chamber is connected to a tank volume. The 3/3-way valve is variably adjustable between these two end positions.

Fig. 1

leine schematische Darstellung eines ersten Ausführungsbeispiels des erfindungsgemäßen hydraulischen Systems am Beispiel einer Einrichtung zur Bremsenergierückgewinnung;

Fig. 2

eine schematische Darstellung zur Verdeutlichung der proportionalen Verstellung des Hubvolumens der hydrostatischen Maschine in Abhängigkeit von einem ersten und einem zweiten Steuersignal;

Fig. 3

eine schematische Darstellung eines abgewandelten Ausführungsbeispiels des erfindungsgemäßen hydraulischen Systems, mit einem durch zwei Proportionalmagnete betätigten Regelventil;

Fig. 4

eine Schnittdarstellung einer beispielhaften Axialkolbenmaschine, die nach dem erfindungsgemäßen System gebaut ist; und

Fig. 5

eine Schnittdarstellung durch ein von zwei Proportionalmagneten betätigtes Regelventil.

A preferred embodiment of the hydrostatic system according to the invention is shown in the drawing and will be explained in more detail in the following description. Show it:

Fig. 1

a schematic representation of a first embodiment of the hydraulic system according to the invention using the example of a device for braking energy recovery;

Fig. 2

a schematic representation to illustrate the proportional adjustment of the stroke volume of the hydrostatic machine in response to a first and a second control signal;

Fig. 3

a schematic representation of a modified embodiment of the hydraulic system according to the invention, with a operated by two proportional solenoids control valve;

Fig. 4

a sectional view of an exemplary axial piston machine, which is constructed according to the system according to the invention; and

Fig. 5

a sectional view through an actuated by two proportional solenoid control valve.

The hydrostatic system 1 according to the invention has, according to a first embodiment, a hydrostatic machine 2. The hydrostatic machine 2 is designed as a pump / motor machine. The hydrostatic machine 2 is designed to be adjustable in terms of its stroke volume and preferably a hydrostatic axial piston machine in swash plate construction, which is pivotable from a neutral position in two directions. The stroke volume is adjusted by adjusting the swash plate. In the illustrated example, the hydrostatic machine 2 is connected to a first working line 3 and a second working line 4. The first working line 3 connects the hydrostatic machine 2 with a hydraulic accumulator 6. The second working line 4, however, is connected to a tank volume 7. Alternatively, it is also possible to provide a high-pressure accumulator as a hydraulic accumulator 6 and a low-pressure accumulator instead of the tank volume 7.

The hydrostatic machine 2 can be operated as a pump, so that it via the second working line. 4 Pressure medium from the tank volume 7 sucks and promotes via the first working line 3 against the pressure prevailing in the hydraulic accumulator 6 pressure. This pressure energy can be stored in the hydraulic accumulator 6. In order to operate the hydrostatic machine 2 as a pump, a drive shaft 5 is provided, which connects the hydrostatic machine 2 with a drive train of a vehicle. By adjusting the delivery volume of the operated as a pump hydrostatic machine 2 with simultaneous connection of the drive shaft 5 to the drive train of a vehicle pressure medium is promoted in a hydraulic brake of the vehicle in the hydraulic accumulator 6 and thus braked the vehicle.

Conversely, an emptying of the hydraulic accumulator 6 via the hydrostatic machine 2 is possible. In this case, the hydrostatic machine 2 is swung in the opposite direction and operated as a motor with a defined set displacement. At the drive shaft 5, an output torque is then generated when emptying the hydraulic accumulator 6 via the hydrostatic machine 2, which is supplied to the acceleration of the vehicle to the drive train. The hydraulic accumulator 6 is emptied into the tank volume 7 via the first working line 3, the hydrostatic machine 2 and the second working line 4.

To adjust the delivery volume or the absorption volume of the hydrostatic machine 2, an adjusting device is provided. The adjusting device comprises a first actuating piston 8 and a second actuating piston 9. The first actuating piston 8 is displaceably arranged in a first actuating cylinder 10 and is connected to the adjusting mechanism of the hydrostatic machine 2 mechanically coupled. By the first control piston 8 thrust forces on, for example, the swash plate as adjustment of the hydrostatic machine 2 can be transmitted. In a corresponding manner, the second actuating piston 9 is arranged displaceably in a second actuating cylinder 11. Also, the second actuating piston 9 can transmit shear forces on the adjusting mechanism of the hydrostatic machine 2, but with opposite effect with respect to the pivoting direction. For this purpose, the adjusting pistons 8 and 9 engage opposite sides of the swashplate of the hydrostatic machine 2, so that a further adjustment of the hydrostatic machine 2 is avoided in the case of an equilibrium of forces between the two forces transmitted to the swashplate by the adjusting pistons 8 and 9.

To generate the actuating forces by the first control piston 8 and the second control piston 9, a first control pressure chamber 12 and in the second control cylinder 11, a second control pressure chamber 13 are formed in the first actuating cylinder 10. The two control pressure chambers 12, 13 are limited by the first control piston 8 or the second control piston 9 on one side. Corresponding to the control pressure prevailing in the first control pressure chamber 12, a hydraulic force is thus generated at a first control piston surface 14 of the first control piston 8 delimiting the first control pressure chamber 12. This hydraulic force is transmitted to the adjusting mechanism of the hydrostatic machine 2 via a corresponding linkage. In the same way, by the pressure prevailing in the second actuating pressure chamber 13 on the second actuating piston surface 15 of the second actuating piston 9 in the opposite direction to the Adjusting mechanism of the hydrostatic machine 2 generates acting force.

In order to set the control pressure prevailing in the first control pressure chamber 12, a control valve 16 is provided. A first connection of the control valve 16 is connected via a control pressure line 17 to the first control pressure chamber 12. Depending on the position of a control element of the control valve 16, the control pressure line 17 is connected to a control pressure source or the tank volume 7. For this purpose, a second connection of the control valve 16 is connected via a connecting line 18 with a shuttle valve 19. With the aid of the shuttle valve 19, either the first working line 3 or else a further pressure medium source is connected via the further connecting line 20 to the connecting line 18 as a control pressure source. Via the further connecting line 20 can be fed from an external pressure medium source pressure medium. This can be useful, for example, if the residual pressure available in the hydraulic accumulator 6 is not sufficient to actuate the adjusting device.

In the connecting line 18, a pressure reducing valve 21 is provided. The pressure reducing valve 21 has a measuring surface which is acted upon by the output pressure of the pressure reducing valve 21, which is supplied to the control valve 16. In the opposite direction acts on the pressure reducing valve 21, the force of a spring over which the start of control of the pressure reducing valve 21 is set. The control valve 16 is thus supplied with a constant input pressure, if on the input side of Pressure Reducing Valve 21 a sufficient pressure is available through the control pressure source.

The control valve 16 has a control valve element, which is preferably designed as a valve piston. This valve piston is arranged axially displaceably in a valve housing and has a first end and a second end remote from it. The first end of the valve piston is acted upon by the control pressure prevailing in the first control pressure chamber 12 and limits the first control pressure chamber 12. Between the first control piston 8 and the valve piston of the control valve 16, a feedback spring 22 is arranged. Depending on the position of the first actuating piston 8, a force acts on the valve piston of the control valve 16, which force is generated by the feedback spring 22 and depends on the position of the first actuating piston 8.

In the opposite direction acts on the valve piston, a counterforce which is generated by a spring 23. The spring 23 is designed to be adjustable, so that the position of the first actuating piston 8, in which the valve piston is in an equilibrium of forces between the feedback spring 22 and the spring 23, can be adjusted. The preferred setting will be described below with reference to FIGS Fig. 2 still explained.

The control valve 16 is designed pressure-compensated. For this purpose branches of the control pressure line 17 from a line branch, which acts on the valve piston at its second end in the same direction with the spring 23 with a hydraulic force. The force acting in the control pressure chamber 12 on the first end of the valve piston hydraulic force is thus by a correspondingly large hydraulic force compensated for the second end. For an integrated arrangement of the control valve 16 in the actuator is possible. The pressure reduction is integrated in the actuator.

In the Fig. 1 the control valve 16 is shown in its neutral position. In this neutral position, the first connection of the control valve 16 is connected in a throttled manner to the second connection of the control valve 16 and to a third connection of the control valve 16. The third port is connected to the tank volume 7. The control valve 16 is designed with negative overlap.

From this neutral position of the control valve 16, the control valve 16 is adjustable in the direction of a first end position and in the direction of a second end position. The control valve 16 can assume any intermediate position between the two end positions. In the first end position of the control valve 16, the first port is connected to the second port of the control valve 16 unthrottled or almost unthrottled. As a result, a connection between the connecting line 18 and the actuating pressure line 17 is generated. In the second end position of the control valve 16, however, the first port and thus the control pressure line 17 is connected to the tank volume 7 via the third port of the control valve 16 unthrottled or almost unthrottled. As a result, the first control pressure chamber 12 is connected in the second end position of the control valve 16 to the tank volume 7 and the pressure in the first control pressure chamber 12 is expanded into the tank volume 7.

The connecting line 18 is also connected in a section between the shuttle valve 19 and the pressure reducing valve 21 to the second control pressure chamber 13. As a result of the actuating pressure prevailing in the second control pressure chamber 13 and the control pressure prevailing in the first control pressure chamber 12, the hydrostatic machine 2 is adjusted until an equilibrium of forces prevails. In this case, the actuating piston surface 15 of the second actuating piston 9 is smaller than the actuating piston surface 14 of the first actuating piston eighth

At the second end of the valve piston of the control valve 16, an actuator 24 is provided. By reducing the pressure by means of the pressure reducing valve 21, the load on the actuator 24 is reduced. This actuator 24 is designed as Doppelhubmagnet with two windings. By applying a first control signal to this first winding, a pressure force can be generated as a control force at the second end of the valve piston. Thus, upon actuation of the actuator 24 by means of the first winding, a control force in a first force direction, which acts in the same direction with the counterforce of the spring 23 on the valve piston, is generated. In addition, in a second opposite direction of force by energizing a second winding of the actuator, a force against the force of the spring 23 can be generated. The resultant of the force of the spring 23 and the force of the actuator 24 acts on the second end of the valve piston.

Is based on the in the Fig. 1 shown rest position of the hydraulic system 1, the actuator 24 is driven in the direction of the first direction of force, the valve piston is adjusted in the direction of its second end position. As a result, the first control pressure chamber 12th relaxed via the control pressure line 17 into the tank volume 7. As a result, the hydraulic force on the first control piston 8 is reduced. Since an unchanged pressure prevails in the second control pressure chamber 13 at the same time, the force on the second control piston 9 is not reduced. Due to the force imbalance, an adjustment of the first actuating piston 8 in the Fig. 1 to the right. As a result, the force of the feedback spring 22 increases to the first end of the valve piston, until again a balance of forces between the force of the feedback spring 22 and the resultant of the control force of the actuator 24 and the opposing force of the spring 23 prevails.

Conversely, by energizing the second winding of the actuator 24, a thrust force against the spring 23 can be generated and the bias of the spring 23 on the valve piston is relieved. Due to the resulting smaller, acting on the second end of the valve piston resultant, the valve piston is adjusted in the direction of its first end position. In the first end position of the control valve 16, the connecting line 18 is connected to the control pressure line 17, so that the first control pressure chamber 12 is supplied from the control pressure source pressure medium. The hydraulic force acting on the first actuating piston 8 is due to the area ratios thus greater than the force acting on the second actuating piston hydraulic force, so that an adjustment of the first actuating piston 8 in the Fig. 1 to the left. This reduces the force generated by the feedback spring 22 on the valve piston at the first end.

The movement of the first actuating piston 8 takes place in each case until, again, a balance of forces between the resultant and the force of the feedback spring 22 has set and the control valve again in the in Fig. 1 shown rest position is. Thus, the position of the first actuating piston 8 and thus the set stroke volume of the hydrostatic machine 2 is in each case proportional to the control signals supplied to the actuator 24 for a respective direction of movement.

The maximum possible pivot angle of the hydrostatic machine 2 is adjusted by mechanical stops which are formed on the adjusting mechanism of the hydrostatic machine 2. In order to bring the hydrostatic machine 2 in a non-pressurized system in a rest position in which it is set to any, in particular to a slightly different from zero delivery volume, a first return spring 25 and a second return spring 26 are provided. With the linkage, which connects the first actuating piston 8 with the adjusting mechanism of the hydrostatic machine 2, a pair of driving elements 27, 27 'is provided. The driving elements 27, 27 'bias the first return spring 25. In an adjustment of the hydrostatic machine 2 by a displacement of the first actuating piston 8 in the Fig. 1 to the right, the first return spring 25 is supported on a displaceably arranged spring plate 28. This is displaceable on the linkage and is based on a movement to the right of the housing side. In an adjusting movement, in which the first actuating piston 8 in the Fig. 1 is adjusted to the right, so that the first return spring 25 is compressed beyond its bias. In the same way, the second return spring 26 is supported via a displaceably arranged further spring plate 29 at an opposite adjustment on the housing side. The second return spring 26 is biased between a pair of further driving elements 30 30 '. Depending on the setting of the hydrostatic machine 2, either the first restoring spring 25 or the second restoring spring 26 is compressed in this case outside of a small adjustment range about the rest position. The respective other return spring 25 and 26 is moved simultaneously with the linkage.

The driving elements 27, 27 'and 30, 30' are arranged on the linkage, that the return springs 25 and 26 are biased and abut the displaceable spring plates 28, 29. The displaceable spring plates 28, 29 are arranged, even if this is not absolutely necessary, in a rest position of the hydrostatic machine 2 at a distance from the housing. In principle, it is also possible without a gap to the housing. Due to the distance component tolerances can be easily compensated. By the bias of the return springs 25, 26 between the driving elements 27, 27 'and 30, 30', the assembly is considerably simplified, since a pre-assembly can take place.

In the Fig. 2 the adjustment of the hydrostatic machine 2 is explained again with reference to a diagram. On the ordinate, the currents I _{1 of} a first control signal and a second control signal I ₂ for energizing the first winding and the second winding are shown. On the abscissa, however, the delivery volume V _P during operation of the hydrostatic machine 2 as a pump or the displacement V _M during operation of the hydrostatic machine 2 are shown as a motor. It can be seen that the hydrostatic machine 2 from an electroless neutral position out in Both directions can be adjusted up to the 100% end position. The spring 23 of the control valve 16 is adjusted so that the neutral position of the control valve 16 is achieved in a position of the first control piston 8, which corresponds to a deflection of the hydrostatic machine 2 from its rest position to a minimum displacement V _{P, min} . Such a setting of the hydrostatic machine 2 with vanishing control signal for the actuator 24 ensures that when the drive shaft 5 is connected to the drive train of a vehicle by the hydrostatic machine 2, a minimum pressure in the first working line 3 by conveying pressure medium into the hydraulic accumulator. 6 is built. This ensures that by the hydraulic system 1 at any time sufficient to actuate the actuator pressure is available and the hydrostatic machine has no intake problems due to wrong pivot side.

The design of the area ratios of the first actuating piston 8 to the surface of the second actuating piston 9 is preferably about 3/1. At the same time, the pressure reducing valve 21 is preferably set to reduce to about 2/3 of the working pressure. The working pressure is the maximum storage pressure of the hydraulic accumulator 6. In the preferred area ratio, a pressure of about 1/3 of the working pressure is required to achieve a force equilibrium between the two adjusting pistons 8,9. By adjusting the pressure reducing valve to 2/3 of the working pressure is thus a sufficient pressure reserve for pivoting the hydrostatic machine 2 in the direction of engine operation available.

Fig. 3 shows a modified embodiment. The function is essentially the same as with reference to Fig. 1 explained function, which is why below only the changes compared to the first embodiment will be discussed. Unlike the in Fig. 1 shown system, the control valve 16 'is not arranged here in the axial extension of the first actuating cylinder 10. The free arrangement of the control valve 16 'allows, as actuating means in place of the Doppelhubmagneten 24, as shown in the Fig. 1 is used to provide a first proportional magnet 24.1 and a second proportional magnet 24.2. The two proportional solenoids 24.1. and 24.2 are each adapted to transmit thrust forces to the valve piston of the control valve 16 '. Since the first actuating piston 8 and the second actuating piston 9 are mechanically coupled to one another via the adjusting mechanism of the hydraulic pump 2, usually a pivoting cradle, the position of the adjusting mechanism can also be tapped via the connection between the second actuating piston 9 and the adjusting mechanism of the hydraulic machine 2 done. This is done in the illustrated embodiment via a linkage 31st

On the linkage 31, the feedback spring 22 'is supported, which acts on the valve piston at one end with the force dependent on the position of the adjusting mechanism. In the opposite direction acts, as already with reference to the Fig. 1 has been explained, a counter force, which is generated by a spring 23 '. With vanishing control force, ie, without applied control signal to the first proportional solenoid 24.1 and the second proportional solenoid 24.2, the force of the feedback spring 22 'and the opposing force of the spring 23' are in an equilibrium of forces when the hydraulic Machine 2 is in its rest position. The deflection of the valve piston of the control valve 16 'now no longer takes place via the Doppelhubmagneten 24, which is disposed on one side of the valve piston, but via either a generation of a thrust force by the first proportional solenoid 24.1 or a thrust force by the second proportional magnet 24.2 on opposite end sides of the valve piston.

Analogous to the above-described function, the hydraulic system also sets itself here so that the resultant of the control force, which is now generated by the two proportional magnets 24.1 and 24.2, and the counterforce 23 'with the force of the feedback spring 22' in equilibrium. In this case, the control valve 16 'in his in the Figure 3 illustrated rest position. If the control signal for the proportional magnet 24.1 or 24.2 is changed, the valve piston is displaced either in the direction of a first end position or the second end position due to the force imbalance. As a result, the control pressure line 17 ', as already with reference to the Fig. 1 has been explained, either connected to the tank volume 7 or the control pressure source. In the illustrated embodiment, the connecting line 18 has no pressure reducing valve. However, this can also in the embodiment according to Fig. 3 be provided.

The arrangement of the second embodiment also has the advantage that only proportional leak pressure can reach the proportional solenoids. A pressure high-pressure resistant construction of the proportional magnets is therefore not required.

Matching elements are provided with matching reference numerals. In order to clarify the similarity of the hydraulic system, the changed elements were provided with primed reference numerals.

The Fig. 4 shows a partial section through an axial piston machine 40, in which the hydraulic system according to the invention finds application. Elements which have already been shown and explained in the schematic representations are designated correspondingly in the sectional representation. Their re-explanation will be omitted to avoid unnecessary repetition. The axial piston machine 40 has a pivoting cradle 41 for adjusting its delivery or absorption volume. This pivoting cradle is tiltable with respect to its inclination angle relative to the axis of rotation of a drive shaft of the axial piston machine 40. To adjust the inclination angle, the pivoting cradle 41, which forms the adjusting mechanism of the hydraulic machine, can be acted upon by thrust forces generated by the first actuating piston 8 and the second actuating piston 9. The position of the pivoting cradle 41 and thus of the set delivery or displacement volume of the axial piston machine 40 is fed back to the control valve 16 by means of the return lever 31. For this purpose, the return lever 31 acts on a movably arranged spring bearing 42. On the spring bearing 42, the feedback spring 22 'is supported, and acts in the manner already described on the control valve 16'. The spring bearing 42 is penetrated by a transmission element 43, by means of which the thrust force generated by the second proportional magnet 24.2 is transmitted to the valve piston of the control valve 16 '. Spring bearing 42 and transmission element 43 do not touch each other.

An enlarged view of the control valve 16 'is in the Fig. 5 shown.

In the section shown, the valve piston 44 of the control valve 16 'can be seen, which is arranged longitudinally displaceable in a valve sleeve. At its first end a spring plate is arranged, on which the feedback spring 22 'is supported. The valve piston 44 is thus acted upon via the spring plate with the force of the feedback spring 22 '. In the opposite direction acts on the valve piston 44, the counterforce generated by the spring 23 '. As already explained, a pair of proportional magnets 24.1 and 24.2 is provided as the actuating means in this embodiment. While the actuating rod 43 of the second proportional magnet 24.2 is supported on the spring plate of the feedback spring 22 'or in parallel with the spring 22' on one side of the valve piston, the plunger of the valve acts on the two opposite ends of the valve piston Proportional magnet 24.1 parallel to the spring 23 'on the opposite end face of the valve piston 44. In the in the Fig. 5 shown rest position of the valve piston 44 is the control pressure line 17 'throttled with the connecting line 18 and connected to the tank volume 7. In the first control pressure chamber 12 therefore acts a mean pressure. For discharging pressure medium in the tank volume holes 45, 46 are provided in the valve sleeve, which communicate with each other and with the spring chamber 47. At the point of implementation of the return arm 38 of the spring chamber 47 with the Housing volume of the axial piston engine 40 connected, which thus forms the tank volume. When the valve piston 44 moves in the axial direction, control edges release an unthrottled or virtually unthrottled connection between the actuating pressure line 17 'and in each case one of two annular spaces formed around the valve piston 44. The annular space oriented closer to the first proportional magnet 24.1 stands in permanent connection with the bore 45 and thus with the tank volume 7. By way of further connection bores, however, the annular space arranged on the side facing the second proportional magnet 24.2 is permanently connected to the connecting line 18.

The counterforce generating spring 23 'is received in a valve piston 44 towards the open sleeve 51, the outside of which is conical in shape and rests against a screwed into the valve sleeve 52 adjusting pin. By differently wide screwing the adjusting pin, the sleeve 51 can be moved axially and thus the force of the spring 23 'can be adjusted. By a lock nut 53, the position of the adjusting pin 52 is secured. Thus, the rest position of the adjusting mechanism is adjustable with vanishing control force.

As an alternative to proportional valves, the control valve can also be actuated hydraulically, wherein different high forces are exerted on the valve piston by different high control pressures.

The invention is not limited to the illustrated embodiment. Rather, individual features of the hydrostatic system according to the invention can be combined with each other in an advantageous manner.

Claims (15)

1. Hydraulic system with an adjustable hydrostatic machine (2) having an adjustment mechanism, with an actuating device which cooperates with the adjustment mechanism and which has at least one actuating piston (8) delimiting an actuating pressure space (12), and with a regulating valve (16, 16') by means of which the feed and discharge of pressure medium into or out of the actuating pressure space (12) can be controlled for the purpose of displacing the actuating piston (8), the regulating valve (16) having a valve piston which is acted upon with a force of a feedback spring (22, 22'), said force being dependent upon the position of the adjustment mechanism, characterized in that the valve piston (44) is additionally acted upon with a counter force acting counter to the force of the feedback spring (22, 22'), and in that actuation means (24, 24.1, 24.2) are provided, by which the valve piston can be acted upon with a control force acting codirectionally with the counter force or with a control force acting contrary to the counter force.
2. Hydraulic system according to Claim 1, characterized in that, to generate the counter force, a spring (23, 23') acting upon the valve piston (44) contrary to the feedback spring (22, 22') is used.
3. Hydraulic system according to Claim 1 or 2, characterized in that the control force can be set continuously.
4. Hydraulic system according to one of Claims 1 to 3, characterized in that the actuation means has a first proportional magnet (24.1) acting upon the valve piston and a second proportional magnet (24.2) acting upon the valve piston in the opposite direction.
5. Hydraulic system according to one of Claims 1 to 3, characterized in that the actuation means has a double lifting magnet (24) with a first winding and with a second winding for generating the control force in each case in one of the opposite force directions.
6. Hydraulic system according to one of Claims 1 to 5, characterized in that the feedback spring (22) is arranged between the actuating piston (8) and the valve piston.
7. Hydraulic system according to one of Claims 1 to 3, characterized in that the hydrostatic machine (2) is set to a feed volume different from zero in the case of a vanishingly low control force and an equilibrium of forces between the counter force and the force dependent upon the position of the adjustment mechanism.
8. Hydraulic system according to one of Claims 1 to 7, characterized in that the actuating device additionally has a further actuating piston (9) which acts upon the adjustment mechanism contrary to the actuating piston (8) which is adjacent to the control pressure space (12).
9. Hydraulic system according to Claim 8, characterized in that the feedback spring (22') is arranged between the further actuating piston (9) and the valve piston (44).
10. Hydraulic system according to one of Claims 1 to 9, characterized in that an actuating pressure source is connected to the regulating valve (16, 16') via a connecting line (18, 18'), and a pressure-reducing valve (21) is arranged in the connecting line (18).
11. Hydraulic system according to Claim 10, characterized in that the actuating pressure source is a working line (3) of the hydrostatic machine (2) or a further pressure medium source (20), the working line (3) or the further pressure medium source (20) being connected to the connecting line (18) via a shuttle valve (19).
12. Hydraulic system according to Claim 10, characterized in that, in a first end position of the regulating valve (16, 16'), the actuating pressure space (12) delimited by the actuating piston surface (14) of the actuating piston (8) is connected to the actuating pressure source.
13. Hydraulic system according to Claim 12, characterized in that, in an opposite second end position of the regulating valve (16, 16'), the first actuating pressure space (12) delimited by the actuating piston surface (14) of the actuating piston (8) is connected to a tank volume (7).
14. Hydraulic system according to Claim 13, characterized in that, in a neutral position of the regulating valve (16, 16') assumed between the first end position and the second end position, the actuating pressure space (12) delimited by the actuating piston surface (14) of the actuating piston (8) is connected, throttled, to the tank volume (7) and, throttled, to the actuating pressure source.
15. Hydraulic system according to a preceding claim, characterized in that the feedback spring (22') is arranged between a transmission part (31), which is fastened to an actuating piston (9), and the valve piston (44).

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