DE102015201318B4 - Hydraulic control arrangement for supplying pressure medium to at least two hydraulic consumers - Google Patents

Abstract

Hydraulic control arrangement for supplying pressure medium to at least two hydraulic consumers (12, 13) with a hydraulic pump (30) whose stroke volume is adjustable, the setting of which can be changed by a pump controller (31) as a function of the highest load pressure of the actuated hydraulic consumers (12, 13) such that the pump pressure is higher than the highest load pressure of the simultaneously actuated hydraulic consumers (12, 13) by a pump pressure differential, with at least two valve arrangements (10, 11), each of which comprises a metering orifice (14, 24) and an individual pressure compensator (16, 26) arranged in series with the metering orifice (14, 24), and of which the first valve arrangement (10) is arranged between a pump line (33) leading from the hydraulic pump (30) and the first hydraulic consumer (12) and the second valve arrangement (11) is arranged between the pump line (33) and the second hydraulic consumer (13), thereby characterized in that the pump controller (31) is controlled in such a way that different pressure drops in the pump line (33) are taken into account in the height of the pump pressure differential and the individual pressure balance (16, 26) assigned to the hydraulic consumer with the highest load pressure (12, 13) is at least almost completely open in the case of different pressure drops in the pump line (33).

Description

The invention relates to a hydraulic control arrangement for supplying pressure medium to at least two hydraulic consumers, comprising a hydraulic pump with an adjustable stroke volume. The pump's stroke volume is adjusted by a pump controller in relation to the highest load pressure of the actuated hydraulic consumers such that the pump pressure exceeds the highest load pressure of the simultaneously actuated hydraulic consumers by a pump pressure differential. At least two valve arrangements are provided, each comprising a metering orifice and an individual pressure compensator arranged in series with the metering orifice. A first valve arrangement is located between a pump line leading from the hydraulic pump and a first hydraulic consumer, and a second valve arrangement is located between the pump line and a second hydraulic consumer. Since the pump pressure is adjusted in relation to the highest load pressure, such hydraulic control arrangements are also referred to as load-sensing control arrangements. Hereinafter, such a control arrangement will be referred to as an LS control arrangement.

Such a hydraulic load-sensing control arrangement is, for example, from the DE 197 14 141 A1 This is a known hydraulic load-sensing control arrangement. In this known hydraulic load-sensing control arrangement, the pressure compensators arranged in series with the metering orifices are acted upon in the opening direction by the pressure downstream of the respective metering orifice and by a spring, and in the closing direction by the pressure upstream of the metering orifice. The pressure compensator is usually arranged upstream of the metering orifice, so that the pressure downstream of the metering orifice is the load pressure of the respective hydraulic consumer.

In another type of hydraulic load-sensing control arrangement, the pressure compensators are located downstream of the metering orifices and are acted upon in the opening direction by the pressure after the respective metering orifice and in the closing direction by a control pressure present in a control chamber, which typically corresponds to the highest load pressure of all hydraulic consumers supplied by the same hydraulic pump. When several hydraulic consumers are actuated simultaneously, if the metering orifices are opened to such an extent that the hydraulic fluid quantity supplied by the hydraulic pump (adjusted to its maximum) is less than the total required hydraulic fluid quantity, the hydraulic fluid quantities flowing to the individual hydraulic consumers are reduced proportionally, regardless of the respective load pressure of the hydraulic consumers. This is therefore referred to as a hydraulic control arrangement with load-independent flow distribution (LUDV control). Because even in an LUDV control arrangement the highest load pressure is sensed and the hydraulic pump generates a pump pressure that is a certain pressure differential above the highest load pressure, an LUDV control arrangement is a special case of a load-sensing control arrangement.

In conventional load-sensing control systems, the flow rate delivered by the hydraulic pump is regulated so that the pump outlet pressure is a constant differential higher than the highest load pressure. A typical differential is, for example, 22 bar. However, for load-sensing flow control via the metering orifice, typically only 10 bar is required. In a load-sensing control system with individual pressure compensators actuated by the pressures upstream and downstream of the metering orifices, this pressure is regulated by the orifices themselves. The remaining 12 bar, the so-called pressure excess, is distributed, depending on the flow rate, between the pressure drop in the pump line and the otherwise unnecessary pressure drop across the pressure compensator assigned to the hydraulic consumer with the highest load pressure. The pressure excess is selected such that, even at maximum flow rate and/or with an unfavorable flow distribution, the desired pressure differential is still available at each metering orifice. Conversely, this means that with a small flow rate associated with a low pressure drop in the pump line and/or a favorable flow rate distribution, the pressure compensator assigned to the hydraulic consumer with the highest load pressure throttles the entire excess pressure. The pressure compensators assigned to the other simultaneously operated hydraulic consumers with lower load pressures must then also throttle from the increased pressure level at the inlet of the respective valve arrangement to a pressure equal to the lower load pressure plus the pressure difference across the metering orifice.

The invention is based on the objective of creating an LS control arrangement that has higher energy efficiency with only minor changes compared to known LS control arrangements.

This task is solved by a hydraulic control arrangement which includes a hydraulic pump whose stroke volume is adjustable, the setting of which is changed by a pump controller depending on the highest load pressure of the actuated hydraulic consumers such that the pump pressure is higher than the highest load pressure of the simultaneously actuated hydraulic consumers by a pump pressure differential, and at least two valves. The device comprises two valve arrangements, each of which includes a metering orifice and an individual pressure compensator arranged in series with the metering orifice. A first valve arrangement is located between a pump line leading from the hydraulic pump and a first hydraulic consumer, and a second valve arrangement is located between the pump line and a second hydraulic consumer. According to the invention, the pump controller is actuated such that different pressure drops in the pump line are taken into account in the magnitude of the pump pressure differential, and the individual pressure compensator assigned to the hydraulic consumer with the highest load pressure is at least almost fully open when there are different pressure drops in the pump line. If only a single hydraulic consumer supplied by the pump is being actuated, then this is naturally also the hydraulic consumer with the highest load pressure.

The basic idea of the invention is therefore that the excess pressure is adjusted according to the expected pressure drops up to the respective metering orifice, so that the throttling by the pressure compensator assigned to the hydraulic consumer with the highest load pressure, i.e., the pressure gradient across this individual pressure compensator, is very low. The pressure gradient across the pressure compensators assigned to the other simultaneously controlled hydraulic consumers is correspondingly reduced, thus achieving high energy efficiency.

Advantageous further developments of a hydraulic control arrangement according to the invention can be found in the dependent claims.

The invention is particularly advantageously implemented in an LS control arrangement in which the at least two individual pressure balances are subjected, in the closing direction, to the pressure upstream of the respective metering orifice and, in the opening direction, to the pressure downstream of the respective metering orifice and a control spring. This is therefore not a LUDV control arrangement with load-pressure-independent flow distribution, but rather the classic LS control arrangement in which, in the event of an insufficient supply, the hydraulic consumer with the highest load pressure no longer receives the desired flow rate and becomes slower compared to hydraulic consumers with lower load pressures. However, the invention can also be advantageously used in an LUDV control arrangement because, for a given opening cross-section of a metering orifice, the flow rate of the hydraulic fluid across it can then be made independent of the pressure drops in the pump line.

The pressure drop in the pump line depends primarily on the flow rate. Therefore, it is possible to use sensors to measure the rotational speed and stroke volume of the hydraulic pump, determine the flow rate in the pump line using these values, and adjust the pump pressure differential accordingly. This ensures that the pressure at the inlet of a valve assembly corresponds to the desired pressure, preventing or minimizing throttling by the individual pressure balance of the hydraulic consumer with the highest load. In the case of a LUDV control system, the pressure differential across the metering orifices is also correct.

The pump pressure differential is advantageously taken from a stored characteristic curve, which contains values for the pump pressure differential as a function of the flow rate. Alternatively, the pump pressure differential can be calculated online, taking into account the physical relationships and parameters.

Variable displacement pumps used in a load-sensing (LS) control arrangement typically employ a control valve as a pump controller. To reduce the pump's displacement, this valve is actuated by the pump pressure; to increase the displacement, it is actuated by the maximum load pressure and a control spring. The differential pressure can be easily modified by adding an electrically controlled actuator to the control valve in either direction. If the actuator acts in conjunction with the pump pressure, the differential pressure decreases as the force exerted by the actuator increases. Conversely, if the actuator acts against the pump pressure, the differential pressure increases as the force exerted by the actuator increases. The main advantage of this design is that, compared to conventional systems, the system continues to operate without functional limitations if the actuator fails.

It is also possible to detect the maximum load pressure using an electrical pressure sensor and to determine a pump pressure from the detected maximum load pressure, the flow rate, and the pump pressure differential associated with the flow rate. The hydraulic pump can then be pressure-controlled and have an electrically controlled pump controller, whereby the determined pump pressure is supplied to the pump controller of the hydraulic pump as the setpoint pump pressure. Advantageously, the setpoint pump pressure is filtered before being supplied to the pump controller. If the pump controller also has a control spring that acts against the pump pressure at the pump controller, the setpoint pump pressure is The value is lower than the pump pressure setpoint by the pressure equivalent of the control spring.

Another advantage is that pump pressure can be requested using electronic actuators without requiring the generation of a hydraulic setpoint pressure signal. This can potentially eliminate the need for hydraulic control lines leading to the hydraulic pump. With multiple hydraulic consumers equipped with multiple pressure sensors, it is possible to evaluate and prioritize individual pressure queries. Furthermore, the changeover valves or check valves typically used to select the highest load pressure across multiple hydraulic consumers can be eliminated.

It is advantageous if a temperature sensor is available, whereby the viscosity of the pressure medium is inferred from the temperature and the pump pressure differential is set taking into account the viscosity of the pressure medium.

It is also advantageous to have pressure sensors that determine the pressure difference across one or more metering orifices. This allows the actual pressure drop occurring in the machine to be identified and, if necessary, corrected. This is particularly beneficial with a LUDV control arrangement.

Finally, other parameters such as volume flow setpoints, volume flow actual values, positions of the metering orifices or the pressure balances in conjunction with pressure information can be included in the calculation of the pump pressure difference.

Two exemplary embodiments of a hydraulic control arrangement according to the invention are shown in the drawings. The invention will now be explained in more detail with reference to these drawings.

• 1 das erste Ausführungsbeispiel, bei dem am Pumpenregler zur vom höchsten Lastdruck und einer Feder ausgeübten Kraft eine veränderliche Magnetkraft addiert wird und
• 2 das zweite Ausführungsbeispiel, bei dem am Pumpenregler zur von einer Feder ausgeübten Kraft eine veränderliche Magnetkraft addiert wird, die einer vom höchsten Lastdruck plus dem veränderlichen Anteil der Pumpendruckdifferenz entspricht.

They show

• 1 the first embodiment in which a variable magnetic force is added to the force exerted on the pump controller by the highest load pressure and a spring and
• 2 The second embodiment, in which a variable magnetic force is added to the force exerted by a spring at the pump controller, which corresponds to the highest load pressure plus the variable component of the pump pressure difference.

The hydraulic control arrangements shown are LS control arrangements with multiple valve assemblies, of which valve assemblies 10 and 11 are shown in the figures. Typically, several valve assemblies are combined into a control block in a monoblock or disc configuration. A hydraulic consumer 12, designed as a hydraulic differential cylinder, is controllable from valve assembly 10, and a hydraulic consumer 13, also designed as a differential cylinder, is controllable from valve assembly 11 with respect to direction and speed.

The valve assembly 10 includes a metering orifice 14, which is typically formed on a control piston (shown only schematically as a rectangle) and whose opening cross-section determines the amount of hydraulic fluid flowing to the hydraulic consumer 12. The control piston is also designed to select which cylinder chamber of the differential cylinder 12 receives the hydraulic fluid flowing through the metering orifice 14 and from which cylinder chamber the hydraulic fluid is displaced and flows to a tank 15. The selection of which cylinder chamber the hydraulic fluid flows to is made downstream of the metering orifice 14, so that in principle one metering orifice 14 is sufficient for both cylinder chambers of the differential cylinder. However, two metering orifices are usually present, arranged between two control chambers of the valve assembly, and one or the other metering orifice is effective depending on the direction of actuation of the control piston.

The control piston, on which the measuring aperture 14 is located, can be continuously adjusted from a central position in two opposite directions, whereby the opening cross-section of the measuring aperture 14 becomes increasingly larger with increasing adjustment.

Upstream of the metering orifice 14, an individual pressure balance 16 is arranged, which is a throttle with an adjustable flow cross-section. Its throttle piston (not shown in detail) is acted upon by the pressure downstream of the metering orifice 14 and a control spring to increase the opening cross-section, and by the pressure upstream of the metering orifice 14 to decrease it. Due to the control via the metering orifice described above, the pressure balance maintains a pressure differential corresponding to the pressure equivalent of the control spring. In this case, this might be, for example, 10 bar.

According to the valve arrangement 10, the valve arrangement 11 has a metering orifice 24, which is usually formed on a control piston shown only schematically as a rectangle and whose opening cross-section allows the pressure medium flowing to the hydraulic consumer 13 to pass through. The quantity is determined. Furthermore, the control piston is designed to select which cylinder chamber of the differential cylinder 13 receives the quantity of pressure medium flowing through the metering orifice, and from which cylinder chamber pressure medium is displaced and flows to the tank 15. The selection of which cylinder chamber the pressure medium flows to is made downstream of the metering orifice 24, so that in principle one metering orifice 24 is sufficient for both cylinder chambers of the differential cylinder 13. However, two metering orifices are usually present, arranged between two control chambers of the valve assembly, and one or the other metering orifice is effective depending on the direction of actuation of the control piston.

The control piston, on which the measuring aperture 24 is located, can be continuously adjusted from a central position in two opposite directions, whereby the opening cross-section of the measuring aperture 24 becomes increasingly larger with increasing adjustment.

Upstream of the metering orifice 24, an individual pressure balance 26 is arranged, which is a throttle with an adjustable flow cross-section. Its throttle piston (not shown in detail) is acted upon by the pressure downstream of the metering orifice 24 and a control spring to increase the opening cross-section, and by the pressure upstream of the metering orifice 24 to decrease it. Due to the control described above via the metering orifice, the pressure balance maintains a pressure differential corresponding to the pressure equivalent of the control spring. In this case, this might be, for example, 10 bar, as with regard to the metering orifice 14 and the individual pressure balance 16.

The two differential cylinders 12 and 13 are supplied with hydraulic fluid by a single hydraulic pump 30. The hydraulic pump 13 has an adjustable stroke volume, where the stroke volume is the amount of hydraulic fluid delivered by the pump during one revolution of its drive shaft. The delivery rate, or flow rate, generated by the hydraulic pump is determined by the instantaneous stroke volume and the rotational speed at which the hydraulic pump is driven.

The hydraulic pump 30 is equipped with a load-sensing control valve 31, which is typically mounted on the pump housing and has a first port that is fluidically connected to a control chamber adjacent to a control piston of the hydraulic pump. A second port of the control valve 31 is fluidically connected to the pressure port of the hydraulic pump. Finally, a third port is connected to the tank. The control valve has a control piston with which the hydraulic fluid flow between the three ports can be controlled. The control piston is acted upon by the pump pressure in the direction of a connection between the first port and the second port on a first surface, and in the direction of a connection between the first port and the third port by a force equal to the force generated by the highest load pressure on a second surface corresponding in size to the first surface, and is also acted upon by a control spring 32. The pressure equivalent of the control spring 32 is, for example, 12 bar. If the force exerted by the pump pressure exceeds the opposing forces, the control chamber is connected to the pressure port of the hydraulic pump. This allows pressurized fluid to flow into the control chamber, thus reducing the stroke volume of the hydraulic pump. If the force exerted by the pump pressure is less than the opposing forces, the control chamber is connected to the tank. This allows, for example, a spring acting on the control piston to displace pressurized fluid from the control chamber to the tank, thereby increasing the stroke volume of the hydraulic pump.

The hydraulic pump 30 is fluidically connected to the valve arrangements 10 and 11 via a pump line 33.

The highest load pressure of all simultaneously operated hydraulic consumers is selected by changeover valves 34 and transmitted via a load sensing line 35 in the embodiment shown. 1 to the control valve 31 and in the embodiment according to 2 reported to a pressure sensor 39.

Without further measures, the pump pressure differential between the pump pressure and the highest load pressure would be constant and would correspond to the pressure equivalent of the control spring 32 of the control valve 31. The pump pressure differential must be designed such that, despite the pressure drop in the pump line up to the respective individual pressure balance, the pressure differential at the metering orifice assigned to the hydraulic consumer with the highest load pressure does not fall below a value that the associated individual pressure balance attempts to regulate. The pressure drop in the pump line increases with the flow rate through the line and is greatest at maximum flow rate. If the pressure drop is, for example, 12 bar, the pump pressure must be 22 bar higher than the highest load pressure to ensure that a pressure differential of 10 bar is still present at the metering orifice of the hydraulic consumer with the highest load pressure. 12 bar is then the so-called pressure surplus.

However, this also means that with small volume flows and lower pressures... Due to the pressure drop in pump line 33, the individual pressure balances must now reduce the pump pressure by a value equal to the excess pressure minus the pressure drop in the pump line. This also applies to the individual pressure balances of the simultaneously operated lower-pressure hydraulic consumers, which of course must reduce the pump pressure anyway.

In the exemplary embodiment according to 1 The control piston of the control valve 31 is acted upon in the direction of a connection of the first port with the second port on a first surface by the pump pressure and in the direction of a connection of the first port with the third port on a second surface corresponding in size to the first surface by the highest load pressure and also by the control spring 32.

To make energy use more efficient, in the first embodiment, the control piston of the control valve 31 is acted upon by a proportionally adjustable electromagnet 36 in addition to the highest load pressure and the control spring 32 in the direction of the connection of the second port with the third port. This is controlled depending on the volume flow in the pump line 33.

To determine the flow rate, the rotational speed at which the hydraulic pump 30 is driven is measured by the speed sensor 40. The stroke volume of the hydraulic pump is measured by a position sensor 41, which, for example, detects the position of the pump's actuator piston. The flow rate is then calculated by multiplying the stroke volume by the rotational speed and, for greater precision, by the volumetric efficiency of the hydraulic pump in an electronic control unit 42. Using a stored characteristic curve 43, the system determines the pump pressure differential required to ensure that, at the given flow rate of the hydraulic pump 30, the pressure at the inlet of the individual pressure compensator assigned to the hydraulic consumer with the highest load pressure is just high enough that the compensator only needs to throttle very little pressure. Ideally, the pressure would be exactly 10 bar above the load pressure.

Based on the determined pump pressure differential, the electromagnet 36 is now activated. If the flow rate is low, a small pump pressure differential of, for example, 12 bar or less is sufficient. The control spring 32 of the control valve 31 may be set to a pressure equivalent of 12 bar, so that the electromagnet 36 is not activated at the low flow rate considered. If it is determined that the current flow rate requires a pump pressure differential of, for example, 16 bar, the electromagnet 36 is energized in such a way that, in addition to the control spring 32, it generates a force equivalent to a pressure of 4 bar. The pump pressure thus increases to 16 bar. The electromagnet is activated accordingly for other flow rates.

An electromagnet could also act on the control piston of the control valve 31 in the same way as the pump pressure. The control spring would then need to have a pressure equivalent of 22 bar. The electromagnet would then be energized at low flow rates to generate a pump pressure only 10 to 12 bar above the maximum load pressure. At very high flow rates, the electromagnet would not be activated, so the pump pressure would be higher than the maximum load pressure by the pressure equivalent of the control spring.

In the exemplary embodiment according to 2 The volume flow rate of the hydraulic pump 30 is the same as in the embodiment shown in the following example. 1 The pump pressure differential is also determined using a stored characteristic curve 43, ensuring that, for the given volume flow of the hydraulic pump 30, the pressure at the inlet of the individual pressure balance assigned to the hydraulic consumer with the highest load pressure is just high enough that the individual pressure balance only has to throttle very little pressure.

In the exemplary embodiment according to 2 The highest load pressure detected by the pressure sensor 39 is fed by an evaluation circuit 44, in which the signal can also be filtered, to a summing point 45. In the summing point 45, a signal corresponding to the pressure difference by which the pump pressure difference at the given volume flow rate should exceed the pressure equivalent of the control spring 32 is added to the electrical signal corresponding to the highest load pressure. A proportionally adjustable electromagnet 46 is controlled according to the added value. Unlike the electromagnet 36 of the embodiment according to... 1 , which only applies a force that is added to the force of the control spring 32 and to the force generated by the highest load pressure, corresponds to the force of the electromagnet 46 of the embodiment according to 2 the force generated by the highest load pressure plus a force necessary to obtain, together with the force exerted by the control spring 32, the desired pump pressure difference at the given volume flow rate.

The maximum load pressure may, for example, be 180 bar. If the volume flow rate is small, a small pump pressure differential of, for example, 12 bar or less is sufficient. The control spring 32 of the control valve 31 may also be, in the embodiment shown, 2 to a pressure equivalent The pressure is set to 12 bar, so that at the small flow rate under consideration, the electromagnet 46 is controlled such that the force it exerts on the control piston of the control valve 31 corresponds to the load pressure of 180 bar. If it is determined that the current flow rate requires a pump pressure differential of, for example, 16 bar, the electromagnet 46 is energized such that the force it exerts corresponds to a force of 184 bar, namely the highest load pressure plus 4 bar, which, together with the 12 bar of the control spring 32, results in a pump pressure differential of 16 bar.

List of reference symbols

10

Valve arrangement

11

Valve arrangement

12

Differential cylinder

13

Differential cylinder

14

Measuring aperture

15

tank

16

Individual pressure scale

24

Measuring aperture

26

Individual pressure scale

30

hydraulic pump

31

Load-sensing control valve

32

Control spring

33

Pump line

34

Changeover valve

35

Load reporting line

36

electromagnet

39

pressure sensor

40

Speed sensor

41

Position sensor

42

electronic control unit

43

Characteristic curve

44

Evaluation circuit

45

Summation point

46

electromagnet

Claims (11)

Hydraulic control arrangement for supplying pressure medium to at least two hydraulic consumers (12, 13) with a hydraulic pump (30) whose stroke volume is adjustable, the setting of which can be changed by a pump controller (31) as a function of the highest load pressure of the actuated hydraulic consumers (12, 13) such that the pump pressure is higher than the highest load pressure of the simultaneously actuated hydraulic consumers (12, 13) by a pump pressure differential, with at least two valve arrangements (10, 11), each of which comprises a metering orifice (14, 24) and an individual pressure compensator (16, 26) arranged in series with the metering orifice (14, 24), and of which the first valve arrangement (10) is arranged between a pump line (33) leading from the hydraulic pump (30) and the first hydraulic consumer (12) and the second valve arrangement (11) is arranged between the pump line (33) and the second hydraulic consumer (13), thereby characterized in that the pump controller (31) is controlled in such a way that different pressure drops in the pump line (33) are taken into account in the height of the pump pressure differential and the individual pressure balance (16, 26) assigned to the hydraulic consumer with the highest load pressure (12, 13) is at least almost completely open in the case of different pressure drops in the pump line (33). Hydraulic control arrangement according to Claim 1 , wherein the at least two individual pressure balances (16, 26) are subjected in the closing direction to the pressure upstream of the respective measuring orifice (14, 24) and in the opening direction to the pressure downstream of the respective measuring orifice (14, 24) and a control spring (32). Hydraulic control arrangement according to Claim 1 or 2 , wherein sensors (40, 41) are provided which detect the rotational speed of the hydraulic pump (30) and the stroke volume of the hydraulic pump (30), wherein the volume flow in the pump line (33) is determined using the rotational speed and the stroke volume and wherein the pump pressure difference is set as a function of the volume flow. Hydraulic control arrangement according to Patent claim 3 , whereby the pump pressure difference is taken from a stored characteristic curve in which values for the pump pressure difference as a function of the volume flow rate are stored. Hydraulic control arrangement according to Patent claim 3 , where the value for the pump pressure difference is calculated online. Hydraulic control arrangement according to one of the Claims 2 until 5 , wherein the pump controller (31) is a control valve which, in order to reduce the stroke volume of the hydraulic pump (30), is controlled by the pump pressure and, in order to increase the stroke volume of the hydraulic pump (30), by the highest load pressure and, in particular, by a control spring (32), and which is in a The two directions of action can additionally be actuated by an electrically controllable actuator (36). Hydraulic control arrangement according to one of the Claim 1 until 5 , wherein the highest load pressure is detected by an electrical pressure sensor (39), wherein a pump pressure is determined from the detected highest load pressure, the volume flow rate and the pump pressure difference associated with the volume flow rate, wherein the hydraulic pump (30) is pressure-controlled and has an electrically controllable pump controller (31) and wherein the determined pump pressure is supplied to the pump controller (31) of the hydraulic pump (30) as the pump pressure setpoint. Hydraulic control arrangement according to Claim 7 , wherein the pump pressure setpoint is supplied to the pump controller (31) in filtered form. Hydraulic control arrangement according to a preceding claim, wherein a temperature sensor is provided, wherein the viscosity of the pressure medium is inferred from the temperature and wherein the pump pressure differential is adjusted taking into account the viscosity of the pressure medium. Hydraulic control arrangement according to a preceding claim, wherein pressure sensors are provided with the aid of which the pressure difference across a metering orifice or several metering orifices is determined. Hydraulic control arrangement according to a preceding claim, wherein further parameters such as volume flow setpoints, volume flow actual values, positions of the metering orifices or the pressure balances in conjunction with pressure information are included in the calculation of the pump pressure difference.