EP2542779B1 - Regulating device and method for controlling the torque of a drive shaft of a hydrostatic machine - Google Patents

Description

The invention relates to a control device and a method for controlling a torque of a drive shaft of a hydrostatic machine.

In a regenerative drive system, a variable in its stroke volume hydrostatic machine is arranged in an open circuit. The hydrostatic machine can drive hydraulic fluid from a tank or a low-pressure accumulator into a high-pressure accumulator during pumping operation via a drive shaft. If energy is required in the regenerative drive system, the hydrostatic machine is operated with pressure medium from the high-pressure accumulator as a motor and the drive shaft is driven. The delivery or displacement volume of the hydrostatic machine should always be adjusted so that given a pressure of the high-pressure accumulator, a predeterminable braking or acceleration torque is present on the drive shaft of the hydrostatic machine.

In the German Offenlegungsschrift DE 10 2006 058 357 A1 a control device for a regenerative drive system is described. To set the torque of the shaft of the hydrostatic machine, the high pressure of the high pressure accumulator is detected by a sensor and given to a control device. The control device calculates from a requested braking torque and the detected high pressure to be adjusted delivery volume of the pump. The stroke volume of the hydrostatic machine is adjusted by means of a control piston of a control device. The flow rates in and out of the actuator are adjusted via a control valve. The control valve is this, so controlled by a control signal that at the hydrostatic machine sets the calculated delivery volume. Such so-called electro-proportional controls which take into account the measured high pressure of the system are generally known.

Such electro-proportional adjustment depending on a measured high pressure and the requested braking torque has the disadvantage that expensive high-pressure sensors must be used. Furthermore, if the high-pressure sensor fails, it is no longer possible to set the requested braking torque on the hydrostatic machine. Therefore, expensive and fail-safe high-pressure sensors or at least one redundant high-pressure sensor must be used.

The object of the invention is to eliminate the disadvantages of the prior art. In particular, it is an object of the invention to provide a control method and a control device for controlling the torque of the drive shaft of the hydrostatic machine to find that do not need a high pressure sensor for operation and / or their functionality is guaranteed even in case of failure of, for example, monitoring the high pressure sensor.

The object is achieved by the method according to claim 1 and by the inventive control device according to claim 11.

The method according to the invention regulates the torque of a drive shaft of a hydrostatic machine. The hydrostatic machine has an adjusting device for adjusting the stroke volume of the hydrostatic machine. The method comprises the following steps: First, a setpoint torque is specified and a set stroke volume of the hydrostatic machine is detected. To regulate the torque of the drive shaft is a flow in or out of the actuator by means regulated by a control valve. The volume flow is regulated on the basis of a force difference between a control force and an opposite to the control force acting on the control valve force acting. The oppositely acting, acting on the control valve force is generated by the voltage applied to the high pressure side of the hydrostatic machine high pressure and counteracts the control force. The size of the control force is set as a function of the detected stroke volume and the predetermined and detected setpoint torque.

The control device according to the invention is basically suitable for controlling a torque of a drive shaft of a hydrostatic machine, wherein the stroke volume of the hydrostatic machine is adjusted by means of a pressure-medium-loaded adjusting device. The control device comprises a control valve for controlling a volume flow into or out of the actuator, e.g. a control pressure chamber of the adjusting device, for adjusting the stroke volume. Furthermore, the control device includes a target torque setting device for setting a target torque and a stroke volume detecting device for detecting a set stroke volume of the hydrostatic machine. The volume flow through the control valve is adjustable with respect to the direction and preferably also the height on the basis of a force difference between a control force and a force acting in the opposite direction to the control valve force. A control surface of one side of the control valve is connected to the high pressure side of the hydrostatic machine to produce the counteracting force. The control device further has a control device which is suitable for specifying the magnitude of the control force as a function of the detected stroke volume and the predetermined setpoint torque.

An advantage of the solution according to the invention is that the stroke volume of the hydrostatic machine automatically to the High pressure, which drives the hydrostatic machine or against which promotes the hydrostatic machine, is adjusted so that the target torque of the drive shaft of the hydrostatic machine remains virtually unchanged. By taking into account the setpoint torque and the set stroke volume in the determination of the control force, the stroke volume is automatically adjusted to the predetermined setpoint torque by feedback of the currently set stroke volume of the hydrostatic machine. In this case, the pressure of the high pressure side by the direct admission of the control valve in the scheme, without having to measure the high pressure by vulnerable sensors. In addition, the target torque is given and the set stroke volume as a control variable easy to capture. In addition, such a control has the advantage that, for example, existing, pressure-controlled pumps can be used by the control according to the invention for torque control.

The dependent claims relate to advantageous developments of the invention.

It is particularly advantageous that the control force is at least partially a hydraulic force. To assist the control force, e.g. a spring force acting in the same direction with the control force on the control valve to simultaneously ensure a defined rest position of the control valve. The control force then consists of an adjustable component and a fixed component, which is generated by the spring. In particular, it is advantageous that the hydraulic force is generated as an adjustable component of the control force by adjusting an opening pressure of a pressure relief valve.

In a preferred embodiment, the control force is generated as a function of the ratio of predetermined setpoint torque to the detected displacement. By such a ratio of given Target torque and the detected stroke volume a negative feedback of desired torque to displacement is achieved. Thereby, when the desired torque is changed, and thus when the control force is changed, it is achieved that the control force is adjusted in the opposite direction as a result of the resulting change in the stroke volume and the feedback of this stroke volume.

It is advantageous to regulate the control force in each case differently in a first operating mode and a second operating mode. The first mode of operation could be, for example, the pump mode and the second mode of operation could be the motor mode of the hydrostatic machine. As a result, the control can respond to different requirements of the control for the first and second operating modes.

Thus, it is particularly advantageous to specify an adjustable component of the control force, for example the hydraulic portion of the control force, in a first operating mode proportional to the ratio of the preset setpoint torque to the detected displacement volume. Accordingly, the control device is suitable for specifying the control force in a first operating mode proportional to the ratio of the preset setpoint torque to the detected displacement. Thus, with an increase in the predetermined setpoint torque in pumping operation and the associated increase in the stroke volume, the control force is reduced again by the negative feedback of the increasing stroke volume. Due to the indirect proportionality of the control force to the detected stroke volume, a negative feedback of the control force is achieved with the set stroke volume.

It is furthermore advantageous that the adjustable component of the control force in turn has two components in a second operating mode. The first fraction is indirectly proportional to the absolute value of the ratio of predetermined target torque to the recorded stroke volume. The second component is dependent on the difference between the setpoint torque and the actual torque of the drive shaft. Preferably, the second proportion is directly proportional to said difference. The control device is suitable for specifying the control force in accordance with the two components. By the first portion is achieved for precontrol that even with swinging through the hydrostatic machine on the Nullhubvolumen (defined here as the neutral position of the hydrostatic machine) to a maximum displacement of the hydrostatic machine in engine operation with a change in the desired torque already in the right direction is adjusted. It is also ensured by the first share continues to feedback the currently set stroke volume of the hydrostatic machine. However, since the first portion allows only a qualitative adjustment of the stroke volume of the hydrostatic machine for a given stroke volume and setpoint torque in engine operation, only the second, superimposed share is controlled to the exact set stroke volume value of the hydrostatic motor.

It is particularly advantageous that the actual torque of the drive shaft of the hydrostatic machine from the displacement and the pressure prevailing on the high pressure side of the hydrostatic machine high pressure is determined. It is also advantageous to estimate the high pressure based on the control force or the adjustable component of the control force, for example, a hydraulic control pressure of the control force. The control device is adapted to determine the actual torque and the high pressure as described. The actual torque can be calculated by the set stroke volume and the prevailing high pressure. Instead of directly measuring the high pressure, it is estimated with the control pressure, which counteracts the high pressure at the control valve. The control pressure is directly proportional to the high pressure, since the control signal always following the high pressure is regulated. In order to cope with the short-term difference between the actual torque and the setpoint torque that arises when the setpoint torque changes, the high pressure is estimated using a filtered control pressure. The control device is adapted to estimate the high pressure by a filtered control pressure. By such a calculation of the control force, it is not necessary to measure the high pressure.

It is also advantageous to additionally detect the high pressure by a simple pressure sensor and to monitor the control signal for specifying the control force on the basis of the detected high pressure. The control device is suitable for monitoring the control force or the control pressure of the hydraulic part of the control force on the basis of a high pressure measured with a pressure sensor.

It is also advantageous that in the first operating mode and in the second operating mode, the direction of attack of the control force and the opposing force on the control valve is reversed. By dispensing with the spring acting on the control pressure, this may e.g. be achieved by a shuttle valve. By such a reversal of the control direction in the second operating mode, i. the engine operation of the hydrostatic machine, the same control system for the control signal can be used in both operating modes. In such an embodiment, the control signal and thus the predetermined control force is always directly proportional to the torque to be adjusted and indirectly proportional to the detected displacement or to their respective absolute amounts.

In a preferred embodiment, the control valve has a first connection connected to a high-pressure side of the hydrostatic machine, a second connection connected to a low-pressure accumulator or tank, and a third connection connected to a control pressure chamber of the actuating device. The Control valve is preferably continuously displaceable between a first, the first terminal connecting to the third terminal position and a second, the second terminal connected to the third terminal position. Such a control valve allows a simple control of the volume flow in and out of the adjusting device on the basis of a force difference.

It is further advantageous that for generating at least one component of the control force, a control surface of the control valve is connected to a control pressure line and this component of the control force is adjustable via the control pressure by the control device. In particular, it is advantageous that the control pressure line for adjusting the control pressure is connected to a pressure limiting valve whose opening pressure can be set by the control device. It is particularly simple and failsafe to adjust the control force at least in part, for example, in addition to the additional component of a spring force, by adjusting a control pressure at one end of the control valve. It is also advantageous that the control pressure line is connected via a throttle directly to the high pressure side of the hydrostatic machine.

It is also advantageous that the control device comprises a control valve having a first and a second control surface and a shuttle valve, wherein in a first position of the shuttle valve, the first control surface of the control valve with the high pressure side of the hydrostatic machine and the second, oppositely acting control surface of the control valve the control pressure line is connected. In a second position of the shuttle valve, the second control surface of the control valve is connected to the high pressure side of the hydrostatic machine and the first control surface of the control valve to the control pressure line. With such a shuttle valve, the flow direction the control with identical force difference between high pressure-dependent force and control force to be reversed. The control device is adapted to bring the shuttle valve in the first operating mode in the first position and in the second operating mode in the second position. This has the advantage that the control algorithm for the control force can be maintained in the second operating mode as well as in the first operating mode. As described above in connection with the first operating mode, the control force to be provided is therefore proportional to the quotient of the setpoint torque and the set displacement.

The control device according to the invention is particularly advantageous for regenerative drive systems that connect the high pressure side of the hydrostatic machine with a high pressure accumulator and in which a durchschwenkbare hydrostatic machine is used. The control device is also advantageous for power regulators which adapt the torque of the drive shaft of a hydrostatic machine to the load pressure.

Fig. 1

einen schematische Darstellung eines hydraulischen Schaltplans eines Ausführungsbeispiels eines regenerativen Bremssystems mit der erfindungsgemäßen Regelungsvorrichtung;

Fig. 2

eine vereinfachte Schnittzeichnung der verstellbaren hydrostatischen Maschine nach dem ersten Ausführungsbeispiel;

Fig. 3

ein Blockschaltbild der Steuervorrichtung des Ausführungsbeispiels;

Fig. 4

ein Diagramm zur Erläuterung der Erzeugung eines Signals zur Steuerdruckerzeugung;

Fig. 5A

einen Zeitverlauf des Soll- und Ist-Drehmoments der Triebwelle der hydrostatischen Maschine, die nach dem erfindungsgemäßen Regelungsverfahren geregelt wird;

Fig. 5B

einen dazugehörenden Zeitverlauf des Hochdrucks und des Steuerdrucks eines Regelventils;

Fig. 5C

einen ebenfalls dazugehörenden Zeitverlauf des Schwenkwinkels der hydrostatischen Maschine; und

Fig. 6

ein Verfahrensablaufdiagramm des erfindungsgemäßen Verfahrens zur Regelung des Drehmoments der Welle der hydrostatischen Maschine.

In the following an embodiment of the invention will be described with reference to the drawing. The drawing shows:

Fig. 1

a schematic representation of a hydraulic circuit diagram of an embodiment of a regenerative braking system with the control device according to the invention;

Fig. 2

a simplified sectional view of the adjustable hydrostatic machine according to the first embodiment;

Fig. 3

a block diagram of the control device of the embodiment;

Fig. 4

a diagram for explaining the generation of a signal for control pressure generation;

Fig. 5A

a time course of the target and actual torque of the drive shaft of the hydrostatic machine, which is controlled by the control method according to the invention;

Fig. 5B

an associated time course of the high pressure and the control pressure of a control valve;

Fig. 5C

an associated time course of the pivoting angle of the hydrostatic machine; and

Fig. 6

a process flow diagram of the method according to the invention for controlling the torque of the shaft of the hydrostatic machine.

Fig. 1 shows a regenerative drive system 1 according to a first embodiment. The regenerative drive system 1 has an axial piston machine 5, a high-pressure accumulator 3 and a control device according to the invention for controlling a torque of a drive shaft 4 of an axial piston machine 5.

The axial piston machine 5 is connected in an open circuit on a low pressure side via a first working line 6 with a tank volume 7 or alternatively with a low pressure accumulator. The axial piston machine 5 is connectable to the high-pressure accumulator 3 on the high-pressure side via a second working line 8. By this arrangement, it is ensured that the second working line 8 is always high-pressure leader in operation and the first working line 6 is always low pressure leader in operation.

An exemplary constructive realization of an axial piston machine unit 2 in Fig. 2 shown. The axial piston machine unit 2 has the adjustable Axial piston machine 5 as a variable in its stroke volume hydrostatic machine and an adjusting device for adjusting the pivot angle of the axial piston machine 5. The adjusting device has in the first embodiment, two adjusting pistons 9 and 10, which are guided in each case in an actuating cylinder 11 and 12 in the longitudinal direction movable.

Fig. 2 shows in the illustrated step, only the second actuating piston 10 and the second actuating cylinder 12. The two adjusting pistons 9 and 10 are coupled to the rotatably mounted in the housing of the hydrostatic machine 5 swash plate 13 of the axial piston machine 5. The position of the actuating pistons 9 and 10 determines the pivot angle of the swash plate 13 and thus the set displacement. The axial piston machine 5 is continuously adjustable from its neutral position (stroke volume = 0) in the positive and negative directions, so that the axial piston machine 5 can be used both in positive pivoting angles in pumping operation and negative pivoting angles in engine operation. In the exemplary embodiment, the swash plate 13 is adjustable from a minimum swivel angle of -18 ° to an absolute maximum same swivel angle of + 18 ° seen. The axial piston machine 5 has no stable zero position and is held by the spring force of a spring 17 in a pressureless state at maximum positive pivot angle.

The first actuating piston 9 and the first actuating cylinder 11 form a first actuating pressure chamber 14 and the second actuating piston 10 and the second actuating cylinder 12 form a second actuating pressure chamber 15. Alternatively, an adjusting device with only one actuating piston, which limits two actuating pressure chambers in one actuating cylinder, could be used. Furthermore, the adjusting device could be designed so that the axial piston machine 5 is in the zero-stroke position or any other rest position in the pressureless state.

The first and second control pressure chambers 14 and 15 of the first embodiment are each acted upon by a first and second control pressure line 23, 16 with a pressure. The second adjusting pressure chamber 15 is permanently connected via the second actuating pressure line 16 to the second working line 8. Therefore, in the second actuating pressure chamber 15 during operation of the axial piston machine 5, the pressure of the high pressure side acts. The second actuating piston 10 is acted upon in addition to a hydraulic force caused by the pressure in the second actuating chamber 15 by the rectified force of the spring 17. As long as the control pressure chambers 14 and 15 are depressurized, the second actuating piston 10 is pushed out in the direction of the second actuating cylinder 12 and thus forces the swash plate 13 in a maximum adjustable positive swivel angle. The cross-section of the first actuating cylinder 11 and thus the pressure-applied surface of the first actuating piston 9 in the first actuating pressure chamber 14 is greater than the cross section of the second actuating cylinder 12. The area ratio is selected so that at the prevailing in operation in the second working line 8 high pressures and at the same pressure in the first actuating pressure chamber 14, the hydraulic force on the first actuating piston 9 in each actuating position is greater than the hydraulic force plus the spring force on the second actuating piston 10 and the axial piston machine 5 is pivoted in the direction of the maximum negative pivot angle.

The axial piston machine unit 2 is connected to a valve block 18 of the control device according to the invention. The valve block 18 has a control valve 19 for controlling the volume flow in and out of the first control pressure chamber 14. The control valve 19 is a 3/2-way valve. A first connection of the control valve 19 is connected via a first supply line 20 to a high-pressure line 21, which is connected to the high-pressure leading second working line 8. The first connection of the control valve 19 is thus with the high pressure side of the axial piston machine 5 and with the High-pressure accumulator 3 connected. The second connection of the control valve 19 is connected via a tank line 22 to the tank volume 7 of the axial piston machine unit 2. The third connection of the control valve 19 is connected via the first control pressure line 23 via a first throttle 24 to the first control pressure chamber 14. The first throttle 24 limits the possible volume flow and thus the adjustment speed. A control valve piston of the control valve 19 can be brought into two end positions. The control valve 19 is continuously adjustable from the first end position to a second end position of the control piston. In the first end position, the first port is connected to the third port of the control valve 19. In this position, the first actuating pressure chamber 14 is connected to the high-pressure leading first working line 8. In the second position, the second port is connected to the third port. In this position, the first actuating pressure chamber 14 is connected to the tank 7 and the first adjusting pressure chamber 14 is expanded into the tank volume.

On a first control surface of the control piston of the control valve 19, a first hydraulic force acts on the second control surface of the control piston in the opposite direction, a second hydraulic force as the first adjustable component of the control force. For this purpose, the first control surface is connected via a second supply line 29 to the high-pressure line 21, so that there always the high pressure of the second working line 8 acts. On the second side of the control valve piston of the control valve 19, the force of an adjustable in its bias spring 25 of the control valve 19 plus the second hydraulic force acts as an adjustable component. The second hydraulic force plus the spring force form the control force acting on the second control surface of the control valve piston. The bias of the spring 25 of the control valve 19 remains constant during operation, so that adjusted during operation, the control force only by the change of the second hydraulic force of the control pressure becomes. If the first hydraulic force is greater than the control force, the control valve 19 goes to the first position. If the control force is greater than the first hydraulic force, the control valve 19 goes to the second position. The transmitted volume flow at given pressure ratios at the ports in the first and second positions depends on the force difference between the first hydraulic force and the control force.

A third supply line 26 connects to generate the second hydraulic force, the second control surface of the control valve piston of the control valve 19 with a control pressure line 27. The control pressure line 27 is connected via a second throttle 28 to the high pressure line 21 and opens at its end remote from it in the tank line 22nd The first and second supply lines 20 and 29 are connected to the high-pressure pilot line 21 upstream of the second throttle 28.

In the control pressure line 27, a pressure relief valve 30 is arranged. On a control surface of the pressure relief valve 30 acts on the control pressure as a first component of the control force-dependent third hydraulic force. In the opposite direction 30, the force of an adjustable in its bias spring 31 of the pressure relief valve 30 acts in the closing direction. If the control pressure in the control pressure line 27 exceeds the opening pressure set by the bias of the spring 31 of the pressure relief valve 30, the pressure limiting valve 30 opens. Thus, the control pressure in the control pressure line 27 upstream of the pressure limiting valve 30 is set as a function of the set opening pressure of the pressure limiting valve 30. This opening pressure is predetermined by the bias of the spring 31 of the pressure relief valve 30 and the opposing force of an electromagnet 32. The opening pressure and thus the set control pressure in the control pressure line 27 can be reduced by increasing the energization of the electromagnet 32.

Furthermore, the first control pressure line 23 between the first throttle 24 and the third port of the control valve 19 via a connecting line 33 to the tank line 22 is connected. The connecting line 33 has a third throttle 34.

In the second working line 8, a check valve 35 for separating the high pressure accumulator 3 from the high pressure line 21, the second setting pressure chamber 15 and the axial piston machine 5 is arranged to prevent leakage. The check valve 35 is to open or close by energizing or non-energizing another solenoid 36 of the check valve 35.

Furthermore, the regenerative drive system 1, a further pressure relief valve 37, a Nachsaugventil 38 and a storage discharge valve 39. The further pressure relief valve 37 opens when a maximum allowable pressure is exceeded by the high pressure in the second working line 8 to the tank volume 7 out. About the Nachsaugventil 38 is said in the case of an empty high-pressure accumulator 3 in engine operation from the tank 7. The storage discharge valve 39 empties the high-pressure accumulator 3 as a result of an electrical discharge signal.

The control device further comprises an electronic control unit 40 as a control device which is connected via a first control connection 41 to the electromagnet 32 of the pressure relief valve 30 and via a second control connection 42 to the further electromagnet 36 of the check valve 35 and via a third control connection 43 to the storage discharge valve 39 is. Furthermore, the control unit 40 is connected via a fourth control connection 44 to a setpoint torque setting device 45, for example an accelerator pedal or a drive lever. The Target torque setting device 45 outputs the torque to be set on the drive shaft 4 as an electrical signal to the control unit 40. The torque to be set on the drive shaft 4 is also referred to below as the setpoint torque. The controller 40 is connected via a fifth control link 46 to a swing angle detector as a stroke volume detecting device. The swivel angle detector detects the set swivel angle a of the swash plate 13 of the axial piston machine 5 and outputs this as an electrical signal to the control unit 40. The swivel angle detector is in Fig. 2 shown and labeled there with "47". The set pivoting angle a is tapped in the illustrated embodiment on the second actuating piston 10. For this purpose, a sensor element 48 of the swivel angle detector 47 is attached to the second control piston 10, which moves with the second control piston 10 in its longitudinal direction. In the movement range of the sensor element 48, a position detecting device 49 is fixedly attached to the housing of the axial piston machine unit 2. The position detection device 49 detects contactless the position of the sensor element 48 and thus the position of the connected to the sensor element 48 second actuating piston 10. The position detection device 49 converts the detected position of the second actuating piston 10 in a set the swivel angle a of the swashplate-indicating signal and sends it over the fifth control connection 46 to the control unit 40. The invention is not limited to the stroke volume detection described here. Rather, any other mechanical, magnetic, electrical or optical Hubvolumenerfassung the axial piston machine 5 is also possible.

Fig. 3 shows a block diagram of the controller 50 of the control device according to the invention. The controller 50 includes the swing angle detector 47, the target torque setting device 45, and the controller 40 on. The controller 40 includes an operation mode detector 51 and first and second control pressure setting sections 52 and 53. The operation mode detector 51 determines whether the axial piston machine 5 is in the pumping mode as the first operating mode or in the engine operating mode as the second operating mode. This can be determined in the first embodiment, for example, from the sign of the swivel angle a or if the swivel angle a is zero from the sign of the setpoint torque T to be set . The operating mode detector 51 is adapted to receive via the input 54 of the control unit 40 a swivel angle α and via the input 55 a signal representing the desired torque T. The operation mode detector 51 is connected to both control pressure command sections 52 and 53 and to a check valve control 62. The operation mode detector 51 is adapted to apply the target torque T and the swivel angle a to the first control pressure command section 52 when the axial piston engine 5 is in pumping operation or to the second control pressure commanding section 53 when the axial piston engine 5 is in engine operation. Further, the operation mode detector 51 is adapted to notify the particular operation mode of the lock valve controller 62, that is, to drive the solenoid 36. The first or second control pressure setting section 52 and 53 calculates a control pressure and converts the calculated control pressure into a control pressure signal that adjusts the calculated control pressure by adjusting the opening pressure at the pressure limiting valve 30. The control pressure signal is applied to the output 56 of the controller 40 from either the first or the second control pressure setting section 52 or 53 and applied to the solenoid 32 via the first control connection 41. The control pressure signal (or the calculated, underlying control pressure) takes into account the proportion of the spring 25 to the control force. For simplicity, the force of the spring 25 is neglected below and only the adjustable component of the control force, so considered the control pressure.

wobei K ₁ eine erste Konstante ist.The first pilot pressure setting portion 52 is active in the determination of the first operating mode by the mode detector 51, and calculates a control pressure p, which is directly proportional to the ratio of the target torque T to pivot angle a. The control pressure p of the first control pressure setting section 52 is calculated to be $$p={\mathrm{<figure-callout\; id="1"\; label="K"\; filenames="imgf0001.png,imgf0003.png"\; state="\{\{state\}\}">K</figure-callout>}}_1\cdot \frac{T}{\alpha }.$$

where K _{1 is} a first constant.

berechnet, wobei K ₂ eine zweite Konstante ist. Die zweite Komponente p ₂ regelt die erste Komponente p ₁ nach. Dazu wird eine Differenz zwischen Soll-Drehmoment T und Ist-Drehmoment T_ist in einem Differenzglied 58 ermittelt, in einem Verstärker 59 verstärkt und zu der ersten Komponente p ₁ addiert. Das Ist-Drehmoment T_ist berechnet sich in der Ist-Drehmomentabschätzung 60 aus dem Schwenkwinkel α und dem die Axialkolbenmaschine 5 antreibenden Hochdruck. Da der Steuerdruck p immer so geregelt wird, dass er dem Hochdruck der Hochdruckleitung 21 etwa entspricht oder zumindest unmittelbar folgt, kann der Hochdruck mit dem vorgegebenen Steuerdruck p = p ₁ + p ₂ abgeschätzt werden. Damit berechnet sich das Ist-Drehmoment T_ist in der Ist-Drehmomentabschätzung 60 zu $${\mathit{T}}_{\mathit{lst}}={\mathit{K}}_3\cdot \mathit{\alpha }\cdot {\mathit{p}}_{\mathit{Hd}}\mathit{.}$$

The second control pressure setting step 53 is active in the second operation mode and in Fig. 4 described in more detail. The output control signal is calculated from a first pressure component p ₁ and a second pressure component p ₂ . In this case, the first component p ₁ is directly proportional to the absolute value of the ratio of the swivel angle α to the setpoint torque T. If the torque T of the drive shaft is to be negative for motor operation and positive for the pumping operation, the absolute value can also be dispensed with in motor operation the swing angle is also negative. The first component p ₁ is in a Vorsignalregelung 57 to $${\mathrm{<figure-callout\; id="1"\; label="p"\; filenames="imgf0001.png,imgf0003.png"\; state="\{\{state\}\}">p</figure-callout>}}_1={\mathrm{<figure-callout\; id="2"\; label="K"\; filenames="imgf0002.png"\; state="\{\{state\}\}">K</figure-callout>}}_2\cdot \left|\frac{\alpha }{T}\right|$$

calculated, where K _{2 is} a second constant. The second component p ₂ regulates the first component p ₁ . For this purpose, a difference between the setpoint torque T and actual torque T _{ist is determined} in a differential element 58, amplified in an amplifier 59 and added to the first component p ₁ . The actual torque T _is calculated in the actual torque estimate 60 from the swivel angle α and the axial piston machine 5 driving high pressure. Since the control pressure p is always regulated so that he High pressure of the high-pressure line 21 corresponds approximately or at least immediately follows, the high pressure with the predetermined control pressure p = p ₁ + p ₂ can be estimated. Thus, the actual torque T _{ist is} calculated in the actual torque estimation 60 $${\mathit{T}}_{\mathit{lst}}={\mathit{<figure-callout\; id="3"\; label="K"\; filenames="imgf0001.png"\; state="\{\{state\}\}">K</figure-callout>}}_3\cdot \mathit{\alpha }\cdot {\mathit{p}}_{\mathit{hd}}\mathit{,}$$

Since the control pressure p deviates from the high pressure briefly in the event of a rapid change in the setpoint torque T , the control pressure p in the correction device 61 undergoes filtering or smoothing before processing in the actual torque estimation 60. The correction device 61 can furthermore or alternatively contain a logic which correctly evaluates the control pressure p in the limiting cases, eg when the setting device is at the stop. The first component p _{1 of} the control pressure p can also be readjusted by more complex controllers, such as a PI controller.

The control unit 40 further comprises the check valve control 62, which energizes the electromagnet 36 of the check valve 35 via the second control connection 42 during operation of the drive system 1 for filling or emptying the high-pressure accumulator 3 in order to connect the high-pressure accumulator 3 to the second work line 8. Furthermore, the check valve control 62 is adapted to close the check valve 35 when the operation mode detector 51 detects neither a pump nor a motor operation and the axial piston machine 5 is set to zero displacement. The control unit 40 also includes an emptying signal transmitter 63, which, for example in the case of maintenance or repair via the third control connection 43, can give a signal to the accumulator discharge valve 39 in order to empty the high-pressure accumulator 3.

The FIGS. 5A, 5B and 5C show an exemplary time course of essential quantities of Control method of the regenerative drive system 1. Fig. 5A shows the time course of the target torque T as a solid curve as it is for example given by a driver and the actual torque T _is as a dashed curve. Fig. 5B shows the pressure-time diagram of the control pressure p as a solid line and the high pressure in the high-pressure accumulator 3 and with open check valve 35 in the second working pressure line 8 as a dashed line. Fig. 5C shows the time course of the set and detected pivot angle α of the axial piston machine. 5

The process according to the invention is carried out on the in Fig. 5 shown, exemplary time course in connection with the method steps for controlling the torque on the drive shaft 4 of the axial piston machine 5 of the drive system 1 in Fig. 6 described.

At the time t ₀ , the axial piston machine 5 is set to the pivoting angle 0 °. This is done, for example, by closing the check valve 35 and by controlling the actuator by a in Fig. 1 realized state machine not shown. Since the check valve 35 is closed, there is no high pressure in the second working line 8. The high-pressure accumulator 3, however, is biased for example to 100 bar.

At the time t ₁ should be braked with the drive system 1. The target torque setting device 45 predefines a brake torque of the drive shaft 4 in a first step S1, ie, a positive target torque. In step S2, the pivot angle of the swash plate 13, which is at about 0 °, is detected. The detected swing angle α and the detected target torque T are supplied to the operation mode detector 51. The operation mode detector 51 detects the operation mode in a third step S3. Since the swivel angle is almost 0 °, the operating mode from the future operating mode, ie from the sign of Target torque T determined. Since the target torque T is positive here in the example, the axial piston machine 5 is to be operated in pumping mode, ie in the first operating mode. In a fourth step S4, the check valve 35 is energized to connect the high-pressure accumulator 3 now with the second working line 8 when a pump or engine operation is detected. Otherwise, in an idle state, when the axial piston machine 5 is at zero displacement or should be adjusted, the energization of the electromagnet 36 of the check valve 35 is interrupted.

As a result of the connection, the high pressure in the second working line 8 rises to the 100 bar of the high pressure accumulator 3. The operation mode detector 51 outputs the target torque T and the pivot angle a to the first control pressure setting section 52 when the operation mode detector 51 detects the first operation mode. As described above, the first control pressure setting section 52 outputs, in a step S5, a control pressure p which is proportional to the ratio of the target torque T and the swing angle α. The control pressure p is initially almost infinitely large due to the finite setpoint torque and almost vanishing pivot angle α in the denominator. The control pressure p is therefore limited and set to a predefined maximum value. The calculated control pressure p is converted into a control pressure signal and given to the output 56 of the controller 40.

In step S7, the control pressure signal is transmitted via the first control connection 41 to the electromagnet 32 of the pressure limiting valve 30, whereby the opening pressure of the pressure limiting valve 30 is set to the calculated maximum control pressure. In step S8, the volume flow in or out of the first control pressure chamber 14 is adapted to the pressure conditions and thus the force difference on both sides of the control valve piston of the control valve 19. By the opening of the check valve 35 is applied to the first side of the control valve piston to the high pressure of the high-pressure accumulator 3. By adjusting the opening pressure of the pressure relief valve 30, a maximum control pressure is applied to the second side of the control piston. Then, the control valve piston is moved from the neutral position in the direction of the second position. The first actuating pressure chamber 14 is connected to the tank volume 7 and hydraulic fluid flows from the first actuating pressure chamber 14 into the tank 7. Therefore, the first actuating piston 9 is pressed into the first actuating cylinder 11 and the axial piston machine 5 is adjusted in the direction of larger positive pivoting angle, ie in the direction of larger delivery volume.

The control steps S1 to S5 or S6 are repeated as long as the control is in operation. A loop duration is significantly shorter than the adjustment time of the swash plate 13 of the axial piston machine 5. Thus, in the first embodiment, the displacement of the axial piston machine 5 is e.g. 100 milliseconds and the loop duration 5 milliseconds. The loop is traversed 20 times within the adjustment time of the axial piston machine 5 from a minimum swivel angle to a maximum swivel angle. The loop duration is an order of magnitude below the adjustment time of the axial piston machine 5. At the same time, in steps S7 and S8, the volume flow in or out of the first control pressure chamber 14 is adapted to the high pressure and the control pressure.

angepasst. Durch den größeren Schwenkwinkel a sinkt der Steuerdruck p bis die Steuerkraft der ersten hydraulischen Kraft entspricht und das Ist-Drehmoment T_ist das Soll-Drehmoment T erreicht hat. Sobald der Schwenkwinkel α positiv ist, fördert die Axialkolbenmaschine 5 Hydraulikflüssigkeit in den Hochdruckspeicher 3 und erhöht so stetig den in der zweiten Arbeitsleitung 8 herrschenden Hochdruck. Der Steuerdruck p wird dem Hochdruck nachgeregelt, so dass ein Gleichgewicht zwischen der ersten hydraulischen Kraft und der Steuerkraft besteht. In dem gezeigten Beispiel wird das vorgegebene Soll-Drehmoment T stetig angehoben, so dass der Steuerdruck p größer als der Hochdruck bleibt.Now that the pivot angle of the axial piston machine 5 is adjusted in the direction of larger positive pivot angle, in steps S1 to S5, the control pressure at the changed pivot angle according to $p={\mathrm{<figure-callout\; id="1"\; label="K"\; filenames="imgf0001.png,imgf0003.png"\; state="\{\{state\}\}">K</figure-callout>}}_1\cdot \frac{T}{\alpha }$

customized. By the larger pivotal angle a of the control pressure p is lowered until the control force of the first hydraulic motor and corresponding to the actual torque T _is the required torque T has reached. As soon as the swivel angle α is positive, the promotes Axial piston machine 5 hydraulic fluid in the high pressure accumulator 3 and thus increases steadily prevailing in the second working line 8 high pressure. The control pressure p is readjusted to the high pressure, so that there is a balance between the first hydraulic force and the control force. In the example shown, the predetermined setpoint torque T is steadily raised, so that the control pressure p remains greater than the high pressure.

At time t ₂ , the setpoint torque T is set constant with about 120 Nm until time t ₃ . The high pressure continues to increase during pumping. Due to the increasing high pressure in the high pressure line 21, the control valve 19 is moved in the direction of the first position and by a pressure medium flow into the first control pressure chamber 14 into the pivot angle α is reduced. Due to the constant adaptation of the swivel angle α to the increasing high pressure and the feedback of the changed swivel angle α, it decreases steadily. By the feedback of the pivot angle α in the calculation of the control pressure p to the control pressure p stabilized below the growing high pressure.

At the time t ₃ , the setpoint torque T is started to increase steadily again. As a result, the swivel angle α of the axial piston machine 5 also increases steadily and the control pressure p is regulated above the high pressure. At the time t ₄ , the axial piston machine 5 moves at the maximum pivot angle α against a stop. The control pressure p increases due to the further increasing setpoint torque T on. Due to the constant maximum pivot angle α, the control pressure p increases in direct proportion to the target torque T and is no longer adapted to the high pressure. Therefore, the control valve piston of the control valve 19 is moved in the direction of the second position and the first actuating pressure chamber 14 connected to the tank 7. The high pressure in the high-pressure accumulator 3 increases now by the constant pumping action on and slowly follows the increasing control pressure p .

At the time t ₅ , the setpoint torque T is kept constant at about 330 Nm until the time t ₇ . The control pressure p stabilizes at about 280 bar and the high pressure continues to increase until the time t ₆ until it reaches the control pressure p . By further increasing the high pressure in the high-pressure accumulator 3 at a constant setpoint torque T , the high pressure line 21 in the high pressure line 21 presses the control valve 19 in the direction of the first position and causes a volume flow in the first control pressure chamber 14. The pivot angle α is smaller, causing the in Dependence of the swivel angle α regulated control pressure p rises again and so follows the high pressure.

After the time t _7, the target torque T is again steadily reduced. As a result, the control pressure p decreases and the swivel angle α reduces further. By the feedback of the sinking pivot angle α, the control pressure p is controlled below the high pressure. At the time t ₈ , the high pressure reaches the opening pressure of the pressure limiting valve 37, which is why the high pressure does not rise further despite the further pumping operation.

From the time t ₉ , the target torque is reduced more slowly. The swivel angle α also travels slower towards 0 °. At the time t ₁₀ , the target torque T reaches a minimum target torque, below which no control of the torque of the drive shaft 4 takes place more and the target torque T is jumped to zero. For very small desired torques T are very small pivot angle α on the axial piston machine 5 and small pivot angle α lead by the inverse relationship to very large control pressures. Therefore, the regulation for small target torques would tend to oscillate. The control is to absolute target torques T greater than one to limiting the minimum nominal torque. Such a minimum nominal torque could also be used to detect a rest position of the axial piston machine 5.

At the time t ₁₀ , the rest position is detected and the check valve 35 is closed. The swivel angle α is held at 0 ° by a state machine.

At the time t ₁₁ , an acceleration torque, ie a negative setpoint torque T , is specified. In steps S1 to S4, the swivel angle α are detected, the setpoint torque T is read in, the operating mode is established and the check valve 35 is opened. The state machine is turned off when a motor or pump operation is detected. The axial piston machine 5 should now be operated in engine operation. Accordingly, the operation mode detector 51 outputs the target torque T and the swivel angle α to the second control pressure commanding section 53, which in conjunction with FIG Fig. 4 has been described. In step S6, the second control pressure generation 53 calculates the control pressure p, and outputs a signal adjusting this control pressure p to the solenoid 32 of the pressure relief valve 30.

As a result of the predetermined acceleration torque, the first portion p _{1 of} the control pressure is zero, since the pivot angle α is still at zero and a second portion of the control pressure greater than zero, since the actual torque is still at zero and a target torque smaller Zero is specified. As a result, the control pressure p decreases and the axial piston machine 5 is swung out in the direction of the negative swivel angle α until the actual torque T _has been adjusted by the setpoint torque T. In order to set the further falling target torque T on the drive shaft 4, the control pressure p is set below the high pressure and drives the pivot angle α further in the direction of a larger displacement volume. By the engine operation it comes to a steady discharge of the high pressure accumulator 3, whereby the axial piston machine 5 driving high pressure decreases.

From the time t ₁₂ , the target torque T is set constant until about the time t ₁₃ at about -150 Nm. Due to the slowly falling high pressure, the control valve 19 is displaced in the direction of the second position and the first actuating pressure chamber 14 is connected to the tank 7. As a result, the swivel angle α changes in the direction of a smaller absorption volume, although the swivel angle α would have to be adjusted in the direction of the maximum absorption volume. By the readjustment of the control pressure p over the second portion p ₂ , this is corrected and the pivoting angle α is changed in the direction of the maximum absorption volume.

From the time t ₁₃ , the amount of the predetermined setpoint torque T decreases again steadily until the time t ₁₄ . As a result, the control pressure p increases through the larger first portion p ₁ and the axial piston machine 5 pivots in the direction of smaller displacement volume, ie in the direction of the neutral position. Due to the smaller swivel angle α, the control pressure p is reduced, so that the control pressure p is set below the high pressure. The deviation between the control pressure p and the high pressure is greater in the engine operation than in the pumping mode, since at the negative adjustment angles the spring 17 is compressed more and thus generates a greater counterforce on the adjusting device.

From the time t ₁₄ , the amount of the desired torque T is increased again and the proportion p ₁ of the control cam and thus the control pressure p decreases. The available high pressure is no longer sufficient to apply the setpoint torque T and the axial piston machine 5 does not adjust the swing angle α fast enough to maximum displacement due to the first throttle 24 and the low high pressure. The amount of the swivel angle α is increased until the axial piston machine 5 reaches the stop at maximum absorption volume at the time t ₁₅ . Then the amount of the actual torque T _ist drops and the control pressure p drops rapidly due to the large difference between the setpoint torque and the actual torque. At the time t ₁₆ , the target torque T is slowly reduced back to zero and the first portion of the control pressure p ₁ increases due to the decreasing absolute target torque T and the second portion of the control pressure p ₂ increases due to the decreasing difference between the target - and actual torque T, T _is . Thereupon, the control pressure p increases, thus retracting the swivel angle a in the direction of zero. When approaching the target torque T of the actual torque T _is at the time t ₁₇ to stabilize the control pressure p and the actual torque _T increases with the target torque T to zero. At the time t ₁₈ , the high-pressure accumulator 3 is emptied and the pressure in the second working line 8 drops suddenly. Due to the lack of pressure in the control pressure line 27 and the high-pressure line 21, the volume flow control of the control valve 19 no longer works and the axial piston machine 5 is held by the already mentioned state machine to a pivot angle a of 0 °. First, the axial piston machine 5 pivots via the spring force 17 via neutral into pumping operation. The then rebuilt pressure allows the use of the state machine.

berechnet werden.The invention is not limited to the embodiment described. The first and the second operating mode could alternatively be realized by the exchange of the attack side of the control force and the first hydraulic force to the control valve 19 by a shuttle valve. As a result, the control pressure p could be described in any case as in the first operating mode of the embodiment $p={\mathrm{<figure-callout\; id="1"\; label="K"\; filenames="imgf0001.png,imgf0003.png"\; state="\{\{state\}\}">K</figure-callout>}}_1\cdot \frac{T}{\alpha }$

be calculated.

The method and the control device according to the invention are not limited to use in regenerative drive systems. Rather, in principle any hydrostatic machine whose torque is to be preset, according to the invention controllable. Thus, the invention is also applicable to power regulators, which are torque regulators per se, when a torque is to be specified. Instead of the pressure in the high-pressure accumulator, the pressure caused by a hydraulic resistance can then occur, such as the load pressure in the case of a load moved by a hydraulic cylinder or a pressure determined by a connected pressure limiting valve. Furthermore, it is also possible to control hydrostatic machines whose high pressure side changes. Here, only in the change, the respective high-pressure leading working line must be connected to a control surface of the control valve 19. Furthermore, the invention can provide only one mode of operation or more than two modes of operation.

In the described embodiment, only the motor and pumping operation for a direction of rotation of the drive shaft 4 and a direction of travel is provided. Thus, the first mode of operation could also involve engine operation in the reverse direction of rotation of the shaft 4 when, for example, the hydrostatic machine is arranged in a closed circuit. In this case, only the high-pressure line 21 must be connected to the corresponding high-pressure working line, so that within the first operating mode still an operating mode case distinction would have to take place, the control of the control pressure p would remain the same. The change between the working lines could be done by a shuttle valve that automatically connects a high pressure side working line with the first side of the control valve 19. Possibly. Signs of the desired torque and the swivel angle must be taken into account so that the control pressure always remains positive. Accordingly, the second mode of operation could also be the Pumping operation in the reverse direction of the drive shaft 4 include. Such an embodiment would have a total of four operating modes, but basically retain the described control.

It would also be conceivable to apply the regulation according to the invention to the deceleration and acceleration of a hydromotor-driven winch, e.g. the catapult-like acceleration and deceleration of the vehicles of a roller coaster with a hydromotor driven winch and a rope.

Furthermore, the invention finds in a start / stop system a particularly advantageous application. Thus, when restarting an internal combustion engine of the regenerative drive system 1, a starting torque can be given to the control device according to the invention and the setting angle of the axial piston machine 5 is automatically set to the correct position for generating the starting torque. Thus, a regenerative hydraulic starter can be realized with the invention particularly simple.

A fan could be driven by an inventively controlled hydraulic motor, which is supplied from a pressure line with pressure medium. The pressure line is supplied by a constant motor and so a certain pressure, which can vary, maintained. At the pressure line could still be connected to a hydraulic accumulator.

The invention is applicable to all hydrostatic machines. The control of the control pressure can be adjusted to each rest position of the actuator in the unpressurized state, such as a central rest position at a Nullhubvolumen.

When applied to the regenerative drive system 1, which is driven by the axial piston machine 5, a disturbance acts, the first change in the regenerative drive system 1 is a change in the hydraulic pressure in the second working line 8. By this pressure change, the hydraulic balance is changed at the control valve 19 and the axial piston machine 5 begins to pivot. This process is recorded electronically and compensated as described. Advantageously, however, the occurrence of the disturbance variable can also be further processed and transmitted at an early point in time as information for a diesel engine as the drive motor of the regenerative drive system 1. This can happen, for example, via a bus connection. Thus, the diesel drive can adjust to the disturbance and it does not have to wait for the dead time to the occurrence of a speed deviation due to the disturbance. This time advantage is particularly advantageous in diesel engine adjustments, which can not react as dynamically due to narrower exhaust limits. A disturbance is, for example, a change in the load torque by changing a gradient to be managed.

The aforementioned embodiments and especially their individual aspects can be combined with each other in an advantageous manner.

Claims (18)

1. Method for regulating a torque of a driveshaft (4) of a hydrostatic machine (5), wherein the hydrostatic machine (5) has an actuating device for setting its piston displacement, the method having the following steps:

   - detecting a predefined setpoint torque (S1, S11);

   - detecting a set piston displacement of the hydrostatic machine (5) (S2, S12);

   - regulating a volume flow into the actuating device or out of the actuating device by means of a regulating valve (19) in order to set the piston displacement on the basis of a difference in the force between a control force and a force (S9, S20) acting in the opposite direction on the regulating valve (19);

   characterized
   in that the force acting on the regulating valve (19) in the opposite direction to the control force is a hydraulic force which is generated by a pressure present on the high pressure side of the hydrostatic machine (5), and which acts counter to the control force, wherein the value of the control force is set (S5, S6; S17) as a function of the detected piston displacement and the detected setpoint torque.
2. Method according to Claim 1,
   characterized
   in that the control force is at least partially a hydraulic force.
3. Method according to Claim 2,
   characterized
   in that the hydraulic force is generated as an adjustable component of the control force by virtue of the setting of an opening pressure of a pressure-limiting valve (30) on the basis of the pressure of the high pressure side (S7, S18).
4. Method according to one of Claims 1 to 3,
   characterized
   in that the control force is generated as a function of the ratio of the detected predefined setpoint torque to the detected piston displacement.
5. Method according to Claim 4,
   characterized
   in that in a first operating mode the control force is proportional to the ratio of the detected setpoint torque to the detected piston displacement (S5, S17).
6. Method according to one of Claims 1 to 5,
   characterized
   in that in a second operating mode the adjustable component of the control force has at least two portions (S6), wherein the first portion is indirectly proportional to the absolute value of the ratio of the detected setpoint torque to the detected piston displacement, and the second portion depends on a difference between the detected predefined setpoint torque and an actual torque of the driveshaft (4) of the hydrostatic machine (5).
7. Method according to Claim 6,
   characterized
   in that the actual torque of the driveshaft (4) of the hydrostatic machine (5) is calculated from a high pressure which is present on the high pressure side of the hydrostatic machine (5) and the detected piston displacement.
8. Method according to Claim 7,
   characterized
   in that the level of the pressure on the high pressure side is estimated on the basis of the control force.
9. Method according to Claim 8,
   characterized
   in that the control force for estimating the high pressure is filtered or smoothed.
10. Method according to one of Claims 1 to 9,
    characterized
    in that in a first operating mode and in a second operating mode the direction of action of the control force and that of the opposing force are interchanged with respect to one another.
11. A regulating device for regulating a torque of a driveshaft (4) of a hydrostatic machine (5) which has an actuating device for setting the piston displacement of the hydrostatic machine (5), the regulating device having a regulating valve (19) for regulating a volume flow into the actuating device or out of the actuating device in order to set the piston displacement on the basis of a difference in force between a control force and a force acting in the opposite direction on the regulating valve (19), a torque-predefining device (45) for predefining a setpoint torque of the driveshaft (4) and a piston displacement-detecting device (46) for detecting a piston displacement which has been set for the hydrostatic machine (5),
    characterized
    in that in order to generate the force acting in the opposite direction on the regulating valve (19) one side of the regulating valve (19) is connected to a high pressure side of the hydraulic machine (5), and in that the regulating device has a control device (40, 66) which is suitable for predefining the value of the control force as a function of the detected piston displacement and the detected setpoint torque.
12. Regulating device according to Claim 11,
    characterized
    in that the regulating valve (19) has a first port which is connected to a high pressure side of the hydrostatic machine (5), a second port which is connected to a low pressure accumulator or tank (7), and a third port which is connected to an actuating pressure chamber (14) of the actuating device, and the regulating valve (19) can be adjusted between a first position which connects the first port to the third port, and a second position which connects the second port to the third port.
13. Regulating device according to Claim 11 or 12,
    
    characterized
    in that in order to set the control force a control surface of the regulating valve (19) is or can be connected to a control pressure line (27), and the control force can be adjusted by means of the level of the control pressure acting there.
14. Regulating device according to Claim 13,
    characterized
    in that the control pressure line (27) has, for setting the control pressure on the basis of the pressure of the high pressure side, a pressure-limiting valve (30) whose opening pressure is adjustable.
15. Regulating device according to one of Claims 11 to 14,
    characterized
    in that a shuttle valve (64) is provided, wherein in a first position of the shuttle valve (64) a control surface of the regulating valve (19) is connected to a high pressure side of the hydrostatic machine (5), and
    a second control surface, acting in opposite fashion, of the regulating valve (19) is connected to the control pressure line (27), and in a second position of the shuttle valve (64) the second control surface of the regulating valve (19) is connected to the high pressure side of the hydrostatic machine (5), and the first control surface of the regulating valve (19) is connected to the control pressure line (27), wherein the control device (66) is suitable for predefining the control pressure of the control pressure line.
16. Regulating device according to one of Claims 11 to 15,
    characterized
    in that the control device (40, 66) is configured to calculate the control force according to one of Claims 4 to 9.
17. Regulating device according to one of Claims 11 to 16,
    characterized
    in that the high pressure side of the hydrostatic machine (5) is connected to a high pressure accumulator (3), and the hydrostatic machine (5) can be adjusted from a neutral position in two opposing directions.
18. Regulating device according to one of Claims 11 to 17,
    characterized
    in that a pressure sensor is provided for detecting the high pressure in order to monitor the regulation of the regulating device.

Images