EP1149246A1

EP1149246A1 - Control arrangement for at least two hydraulic consumers and pressure differential valve for said control arrangement

Info

Publication number: EP1149246A1
Application number: EP00906203A
Authority: EP
Inventors: Martin OBERHÄUSSER; Thomas Weickert
Original assignee: Mannesmann Rexroth AG
Current assignee: Bosch Rexroth AG
Priority date: 1999-02-05
Filing date: 2000-01-15
Publication date: 2001-10-31
Anticipated expiration: 2020-01-15
Also published as: US6644025B1; DE19904616A1; EP1149246B1; WO2000046513A1; JP4739529B2; DE50009158D1; JP2002536599A

Abstract

The invention relates to a control arrangement which is used to supply pressure means to at least two hydraulic consumers and which has a demand-regulated variable displacement pump which can be regulated by a pump regulator in dependence on the highest load pressure of the hydraulic consumers being actuated; two adjustable proportioning diaphragms, the first of which is located between a supply line leading from the variable displacement pump and a first hydraulic consumer and the second of which is located between the supply line and a second hydraulic consumer; and two pressure balances, the first of which is connected downstream of the first proportioning diaphragm and the second of which is connected downstream of the second proportioning diaphragm and whose regulating piston is subjected to the pressure in the opening direction after the corresponding proportioning diaphragm. In order to prevent a temporary surplus of the variable displacement pump from passing through to the hydraulic consumers with a control arrangement of this type, the regulating pistons of the pressure balances can be subjected to a control pressure which is available in a reverse control chamber, in the closing direction. The control pressure is derived from the supply pressure present in the supply line by means of a valve device and changes with the supply pressure. The invention also relates to a pressure differential valve with a small construction which ensures that an increased pressure at its entrance is followed by a pressure at its exit with a fixed pressure difference. Together with a restricted relief of the output to the tank, a pressure differential valve of this type ensures that there is always a fixed pressure difference between the output pressure and the input pressure. A valve of this type is especially suitable for use in the inventive control arrangement.

Description

description

Control arrangement for at least two hydraulic consumers and pressure differential valve therefor

5

The invention relates to a control arrangement with which at least two hydraulic consumers are supplied with pressure medium and which has the features from the preamble of patent claim 1. The invention also relates to a pressure differential valve which is used in particular in the control arrangement mentioned.

A hydraulic control arrangement according to the preamble of claim 1 is known for example from EP 0 566 449 A1. It is a hydraulic control arrangement based on the load-sensing principle, with i5 in which a variable pump is set depending on the highest load pressure of the hydraulic consumers operated so that the inlet pressure is a certain pressure difference above the highest load pressure. The pressure medium flows to the two hydraulic consumers via two adjustable metering orifices, of which a first is arranged between a pump line coming from the variable displacement pump and a first hydraulic consumer and the second between the pump line and the second hydraulic consumer. The pressure compensators connected downstream of the metering orifices ensure that with a sufficient quantity of pressure medium supplied, regardless of the load pressures of the hydraulic consumers, there is a certain pressure difference 5 across the metering orifices, so that the quantity of pressure medium flowing to a hydraulic consumer only depends on the opening cross section of the respective metering orifice. If a metering orifice is opened further, more pressure medium has to flow over it in order to generate the specific pressure difference. The variable pump is adjusted so that it delivers the required amount of pressure medium. One therefore speaks of a demand current control.

The pressure compensators downstream of the metering orifices 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 rear control chamber, which usually corresponds to the highest load pressure of all hydraulic consumers supplied by the same hydraulic pump. If, with simultaneous actuation of several hydraulic consumers, the metering orifices are opened so far that the quantity of pressure medium supplied by the hydraulic pump adjusted to the stop is smaller than the total quantity of pressure medium required, the quantity of pressure medium flowing to the individual hydraulic consumers becomes independent of the respective load pressure of the hydraulic Consumers reduced proportionately. One therefore speaks of a control with load-independent flow distribution (LUDV control). Hydraulic consumers controlled in this way are called LUDV consumers for short. Because the LUDV control also senses the highest load pressure and an inlet pressure that is a certain pressure difference above the highest load pressure is generated by the pressure medium source, a LUDV control is a special case of a load-sensing or load-sensing control (LS control) .

For several hydraulic consumers, to which pressure medium flows via a metering orifice with an upstream pressure compensator, which is only acted upon by the pressure in front of the metering orifice in the closing direction and only by the load pressure of the respective hydraulic consumer and by a compression spring in the opening direction, no load-independent flow distribution is obtained. You have a mere LS control and a LS consumer. Such a control is known for example from DE 197 14 141 A1. With simultaneous actuation of several hydraulic consumers and insufficiently supplied by the variable pump The quantity of pressure medium is reduced here only to the quantity of pressure medium flowing to the hydraulic consumer with the highest load pressure.

An advantage of a LS control with the pressure compensators connected upstream of the metering orifices compared to an LS control with the pressure compensators connected downstream of the metering orifices is, however, that the upstream pressure compensators by reducing the oversupply supplied for a short time by the variable displacement pump and an associated increase in the inlet pressure their opening cross-section do not allow the pressure difference across the metering orifices to increase, so that no more pressure medium flows through the metering orifices and the speed of the hydraulic consumers is not changed. The excess flows back to a tank via a pressure relief valve. In contrast, in the case of a control with the pressure compensators connected downstream of the metering orifices, the excess quantity is passed on to the hydraulic consumers.

Depending on whether the user attaches more importance to a load-independent flow distribution or to a prevention of excessive amounts flowing to the hydraulic consumers, he desires an LUDV control or an LS control. So far, this has been disadvantageous for the manufacturers of hydraulic components, since they have to offer control blocks for both LUDV controls and LS controls. These are very different, since depending on whether a pressure compensator is connected upstream or downstream of the corresponding metering orifice, very different designs are necessary.

In contrast, the invention is based on the objective of designing a hydraulic control arrangement which has the features from the preamble of patent claim 1, in which, in particular, the pressure compensators are connected downstream of the metering orifices, in such a way that the inflow of excess quantities to the hydraulic consumers is prevented. The desired aim is achieved according to the invention in that, in a generic hydraulic control arrangement according to the characterizing part of patent claim 1, the control pistons of the pressure compensators can be acted upon in the closing direction by a control pressure present in a rear control chamber which is derived from the supply pressure prevailing in the supply line with the aid of a valve device is and changes with the inlet pressure. While in the known hydraulic control arrangement with the metering orifices connected downstream, the highest load pressure is exerted on them in the rear control chamber, on which the delivery rate of the variable pump has no influence, in a control arrangement according to the invention the control pressure in the rear control chamber is derived from the inlet pressure and changes with it this. If the inlet pressure rises due to a delivery volume of the variable displacement pump that exceeds the demand, the control pressure also rises. Accordingly, the control pistons of the pressure compensators in

Closing direction moves, so that the pressure after the metering orifices increases and the pressure difference across the metering orifices does not change. A constant pressure difference across an orifice plate means, however, with a constant opening cross section of the orifice plate, also means a constant amount of pressure medium flowing through the orifice plate. Thus, while maintaining the basic arrangement of metering orifice and downstream pressure compensator and thus without fundamental changes to a control block with minor modifications, the same control behavior is achieved as when controlling with the metering orifices upstream of pressure compensators and thus completely different control blocks.

Advantageous embodiments of a hydraulic control arrangement according to the invention can be found in subclaims 2 to 9. Thus, according to claim 2, the difference between the inlet pressure and the control pressure when the variable displacement pump has not yet been adjusted to the stop, that is to say with a sufficient amount of pressure medium, is not greater than between the inlet pressure and the highest load pressure. If the pressure difference were greater, the amount of pressure medium flowing to a hydraulic consumer would depend on whether the load pressure of this hydraulic consumer is higher or lower than the control pressure. The control pressure is preferably slightly higher than the highest load pressure, so that on the one hand there are no unnecessary throttling losses on the pressure compensators, but on the other hand the pressure compensator assigned to the respective hydraulic consumer with the highest load pressure is still in the control range.

Basically, it is conceivable to generate the pressure difference between the inlet line and a rear control chamber on a pressure compensator by connecting a nozzle between the inlet line and the control chamber and a flow control valve between the control chamber and a tank. A certain amount of control oil would flow out of the control room to the tank via the flow control valve. This amount of control oil would flow to the control room via the nozzle. There would thus be a constant pressure drop across the nozzle. However, the amount of pressure medium flowing through a nozzle is strongly dependent on the viscosity of the pressure medium. It therefore appears cheaper to use a pressure difference valve instead of a nozzle according to claim 3, which is connected with an inlet to the inlet line and with an outlet to the rear control chamber of a pressure compensator. The pressure difference valve is preferably set to a fixed pressure difference and has a movable valve member which, in the sense of opening the fluidic connection between the inlet line and the control chamber on the pressure compensator, from the inlet pressure and in the sense of closing this connection from the control pressure and a spring is applied. A particularly preferred embodiment also contains claim 5, according to which the rear control rooms of several pressure compensators are directly connected to one another, so that the same control pressure prevails in these control rooms. Only one valve device for deriving the control pressure from the inlet pressure is therefore necessary for these pressure compensators. In the particularly advantageous embodiment according to claim 6, the control arrangement has a load signal line, into which the highest load pressure of the hydraulic consumers actuated is given via selection valves, and a valve which opens a fluidic connection from the load signal line to the rear control chamber of at least one pressure compensator when the difference between the inlet pressure and the highest load pressure falls below a certain value. In this way, in the event of undersaturation, that is to say if the variable displacement pump does not deliver sufficient pressure medium, a load-independent flow distribution is obtained between the hydraulic consumers, the pressure balances of which are connected with their control chamber to the load signal line.

If a hydraulic consumer is to be supplied primarily with pressure medium compared to another hydraulic consumer in the event of undersaturation, this is advantageously done by an embodiment according to claim 8. The rear control chamber on the pressure balance of the hydraulic consumer which is primarily to be supplied with pressure medium is then separated from the control rooms to the pressure balances of the other hydraulic consumers. The control pressure in it is derived from the inlet pressure via a further valve device. There is also a priority valve which is used to maintain a desired pressure difference across the metering orifice arranged upstream of the pressure balance of the privileged hydraulic consumer and thus to maintain an adequate supply of pressure medium to the privileged hydraulic consumer when the hydraulic the corresponding delivery rate of the variable pump, the control pressure in the rear control chamber of the other hydraulic consumers can be raised above the control pressure in the event of saturation. Preferably, the priority valve according to claim 9 has a first connection connected to the supply line and a second connection connected to the rear control chambers of the pressure compensators assigned to the non-privileged hydraulic consumers and has a valve member which opens in the direction of the connection between the first connection and the second connection from in a line section downstream of the preferred hydraulic consumer

I O associated metering orifice prevailing pressure and an additional force and in the direction of closing the connection between the first connection and the second connection can be acted upon by the inlet pressure. A control chamber of the priority valve can be connected upstream or downstream of the pressure compensator to the line section downstream of the metering orifice, since the priority valve is then in

15 The function occurs when the pressure compensator is completely open and because then the pressure in front of and behind the pressure compensator is the same, namely the load pressure of the privileged hydraulic consumer.

It is also an object of the invention to provide a pressure differential valve which is used in particular to derive a control pressure for a pressure compensator from the inlet pressure in a control arrangement according to one of Claims 1 to 9 and which is particularly small so that it can easily be used in a control block.

25 Such a pressure differential valve is obtained by the features contained in the characterizing part of claim 10.

Advantageous embodiments of such a pressure differential valve are contained in claims 11 to 13. An exemplary embodiment of a control arrangement according to the invention and a pressure differential valve used therein are shown in the drawings. The invention will now be explained in more detail with reference to the figures in the drawings.

Show it

1 shows a circuit diagram of the exemplary embodiment of the control arrangement, which shows LUDV behavior in the event of undersaturation and which contains a privileged hydraulic consumer,

FIG. 1a shows an alternative to actuating the priority valve shown in FIG. 1,

2 shows the circuit diagram of a variable displacement pump used in the exemplary embodiment, and FIG. 3 shows a longitudinal section through the pressure differential valve used in the exemplary embodiment according to FIG.

According to FIG. 1, a variable displacement pump 10 with a variable displacement 11 sucks pressure medium from a tank and releases it into a system of supply lines 13. In the exemplary embodiment, three hydraulic consumers 14, 15 and 16, all of which are designed as differential cylinders, are supplied with pressure medium via the feed lines. To control the speed and the direction of movement, each differential cylinder 14, 15 and 16 is assigned a metering orifice 17, 18 and 19 and a 4/3-way valve 20, 21 and 22, respectively. In practice, a metering orifice and a directional control valve are each integrated in such a way that the direction of movement of the differential cylinder is predetermined by the actuation of a valve spool that is spring-centered in a central position in a certain direction, and the opening cross section is determined by the path that the valve spool is moved the metering orifice is determined. For a concrete constructive solution solution is referred to EP 0 566 449 A1 already mentioned. The metering orifices 17, 18 and 19 are connected to the system of the feed lines 13. Between a metering orifice 17, 18 and 19 and a directional control valve 20, 21 and 22, respectively, a pressure compensator 23, 24 and 25 is arranged, the control piston of which is not shown in the opening direction from the pressure downstream of the respective metering orifice and in the closing direction of a control pressure prevailing in a rear control chamber 26 is applied. The directional control valves 20, 21 and 22 each have two consumer connections 30 and 31 connected to pressure chambers of the respective differential cylinder, an inlet connection 32 which is connected to the outlet of the respective pressure compensator, and a return connection 33, from which a return line leads to the tank 12. In the middle position of a directional control valve, the two consumer connections are shut off and the inlet connection is connected to the tank connection. The line section between the outlet of the pressure compensator and the inlet connection is therefore relieved of pressure. In a lateral working position of a directional valve, pressure medium flows to one pressure chamber of a hydraulic cylinder, while pressure medium can flow away from the other pressure chamber to the tank.

The control pistons of the pressure compensators 23, 24 and 25 are acted upon in the direction of closing in addition to a control pressure by a weak compression spring 34, which has a pressure of e.g. is equivalent to only 0.5 bar. In addition, the control rooms 26 and 27 of the two pressure compensators 23 and 24 are connected to one another via a channel 35, so that the same control pressure is always present in the two control rooms 26 and 27.

Change-over valves 36 are connected to the outputs of the pressure compensators 23, 24 and 25 or to the inlet connections 32 of the directional control valves, which are linked together in such a way that the highest load pressure of all actuated in a load signaling line 37, which leads to the adjustment 11 of the pump 10 Differential cylinder is present. In particular, as can be seen from FIG. 2, the load signaling line 37 leads to a control valve 39 with three connections, one of which is connected to an actuating cylinder 40 of the variable displacement pump 10. Another connection of the control valve 39 is connected to a feed line 13 and the third connection to tank 12. 5 The control piston of the control valve 39 is acted upon in the direction of a connection of the first connection to the second connection by the pressure in the feed line 13 and in the direction of a connection of the first connection to the third connection by the pressure in the load signaling line 37 and by a control spring 41. Variable pumps and control valves according to the circuit diagram according to FIG. 2 are generally known and readily available on the market. It is therefore unnecessary to go into this in more detail. It should only be pointed out that the load-sensing or load-sensing pump control shown has the effect that a pressure is established in the feed line 13 which is a pressure difference equivalent to the force of the control spring 41 above the pressure in the load signaling line 37.

15

A pressure differential valve 45 is arranged between the system of the feed lines 13 and the channel 35 between the two control rooms 26 of the pressure compensators 23 and 24. This is connected to the inlet lines 13 with an inlet opening 46 and to the channel 35 with an outlet opening 47. Depending on the 0 position of a spool 48 of the pressure difference valve 45, which is not visible in FIG. 1 but is shown in FIG. 3, the inlet opening 46 and the outlet opening 47 are blocked off from one another or are fluidly connected to one another via a more or less large opening cross section. The piston slide 48 is in the direction of reducing the opening cross section between the inlet opening and the outlet opening from the pressure prevailing in the channel 35 and in the control chambers 26 of the pressure compensators and from a compression spring 49 and in the direction of increasing the opening cross section from the inlet pressure prevailing in the feed lines 13 acted upon. The effective areas on the spool for attacking the control pressure and the inlet pressure are equal in size, so that the pressure differential valve 45 ensures that the control pressure present in the channel 35 follows an increasing inlet pressure at a distance of a differential pressure equivalent to the force of the compression spring 49. For example, the pressure differential valve 45 is set so that the control pressure is 20 bar lower than the inlet pressure. The channel 35 is connected to the tank 12 via a small flow regulator 50, so that the control pressure in the channel 35 can also follow a decreasing inlet pressure through the discharge of pressure medium via the small flow regulator 50.

A check valve 51 is connected between the load signaling line 37 and the channel 35 and opens from the load-reducing line 37 to the channel 35 when the pressure in the channel 35 becomes equal to the pressure in the load signaling channel 37. The control pressure present in the control rooms 26 of the pressure compensators 23 and 24 cannot therefore fall below the highest load pressure present in the load reporting line 37.

There is a second pressure difference valve 52, which is identical to the pressure difference valve 45 and whose inlet opening 46 is also connected to a feed line 13. The outlet opening 47 of the pressure differential valve 52 is connected to the control chamber 26 of the pressure compensator 25. The control of the spool of the pressure difference valve 52 is carried out in the same way as the control of the spool of the pressure difference valve 45. Both valves are set to the same pressure difference of, for example, 20 bar. With a sufficient delivery rate of the variable displacement pump 10, the control pressure in the control rooms 26 is 20 bar lower than the inlet pressure and, since this should be 25 bar higher than the highest load pressure, 5 bar higher than the highest load pressure. The pressure compensators 23, 24 and 25 are therefore all in control, including those assigned to the consumer with the highest load pressure. Position. Furthermore, the control chamber 26 of the pressure compensator 25 is connected to the tank 12 via a second small current regulator 50.

The differential cylinder 16 should, if the variable pump 10 brings maximum delivery rate and this does not meet the demand, be supplied with pressure medium primarily before the other two hydraulic cylinders 14 and 15. For this purpose, a priority valve 55 is provided, which is designed as a proportional orifice with an inlet 56 and with an outlet 57. The latter is fluidly connected to the channel 35. The inlet 56 is connected upstream of the metering orifice 19 to a feed line 13. The movable valve member of the priority valve, not shown, is acted upon in the direction of closing the connection between the inlet and the outlet by the pressure in the inlet, that is to say by the inlet pressure and in the direction of opening the connection by the pressure downstream of the metering orifice 19 and by the force of a control spring 58. The control spring 58 is designed, for example, in such a way that a force equilibrium exists on the valve member of the priority valve if the pressure difference between the inlet pressure and the pressure downstream of the metering orifice 19 is 19 bar. This value is slightly less than the value of the pressure difference across the pressure difference valve 52 reduced by a pressure value of 0.5 bar equivalent to the force of the compression spring 34. So while in normal operation there is a pressure difference of 19.5 bar via the orifice 19, the priority valve 55 does not respond. If, by reducing the inlet pressure, the pressure difference across the metering orifice 19 drops below 19.5 bar, the pressure compensator 25 opens completely, so that the pressure downstream of the metering orifice 19 is equal to the load pressure of the primary hydraulic consumer 16. The load pressure of the consumer 16 is now present on the priority valve 55 on the spring side. He can open the priority valve 55 against the inlet pressure, whereby the pressure in the channel 35 and thus in the control rooms 26 of the pressure compensators 23 and 24 is raised above the highest load pressure. The pressure compensators 23 and 24 are therefore adjusted in the closing direction until an increase in the pressure downstream of the metering orifices 17 and 18 an equilibrium of forces is reached on their control piston. However, the pressure difference across the metering orifices 17 and 18 is now reduced. The pressure medium flows flowing to the consumers 14 and 15 have become smaller. Ultimately, the priority valve 55 ensures by increasing the pressure in the 5 control rooms 26 of the pressure compensators 23 and 24 that by increasing the control pressure in the channel 35, the pressure difference across the metering orifices 17 and 18 and thus that flowing to the hydraulic consumers 14 and 15 Pressure medium flows are each reduced to such an extent that a quantity of pressure medium flows via the metering orifice 19, which produces a pressure difference which is approximately equal to the IO pressure difference in normal operation.

As already mentioned, in the case of undersaturation, ie if the priority valve 55 is to respond, there is load pressure downstream of the metering orifice 19. Alternatively, therefore, the spring-side control chamber of the priority valve 55 instead of the

15 connection between the metering orifice 19 and the pressure compensator 25 can also be connected to the outlet of the pressure compensator 25, as shown in FIG. 1a. The valve member of the priority valve 55 is then always acted upon by the load pressure of the priority hydraulic consumer 16 in the direction of opening the connection between the inlet 56 and the outlet 57. The 0 priority valve can now be set to the same pressure difference that also prevails in normal operation via the metering orifice 19, since in normal operation the pressure difference between the load pressure of the priority hydraulic consumer 16 and the inlet pressure is higher than the pressure difference via the metering orifice 19 and therefore the priority valve 55 certainly does not respond. 5

If the case of undersaturation occurs when only the hydraulic consumers 14 and 15 are actuated, the control pressure in the channel 35 becomes equal to the highest load pressure of the two hydraulic consumers 14 and 15 in the load signaling line 37 due to the drop in the inlet pressure the check valve 51 the highest load pressure also reported in the channel 35. A further drop in the inlet pressure therefore no longer results in a further drop in the control pressure in the channel 35 and in the control spaces 26 of the pressure compensators 23 and 24. These ensure that a pressure is present between them and the metering orifices 17 and 18, regardless of the level of the inlet pressure, which is higher than the highest load pressure by the pressure equivalent of the springs 34. This pressure, which is slightly higher than the highest load pressure, is present downstream of both metering orifices 17 and 18. Upstream of both metering orifices 17 and 18 there is inlet pressure. Thus, the pressure difference across the metering orifice 17 is equal to the pressure difference across the metering orifice 18. The pressure medium flows to the hydraulic consumers 14 and 15 are therefore reduced proportionately in the event of undersaturation regardless of whether the privileged consumer 16 is actuated. Consumers 14 and 15 are therefore LUDV consumers.

If the pressure medium requirement of all simultaneously operated hydraulic consumers is covered by the variable pump 10, the pressure differential valves 45 and 52 together with the flow regulators 50 ensure that the control pressures in the control chambers 26 of the pressure compensators follow the inlet pressure at a fixed distance. If the variable displacement pump 10 now briefly produces an amount that exceeds what is required, for example because a wide-open metering orifice is completely closed, the inlet pressure rises sharply for a short time. The control pressures follow this increase, so that the control pistons of the pressure compensators are subjected to an increased control pressure in the closing direction, move in the closing direction of the pressure compensators and thereby raise the pressure downstream of the metering orifices, so that the pressure difference across the metering orifices 17, 18 and 19 remains the same or increases only slightly. So the speed of a hydraulic consumer does not increase. The excess flows through a pressure relief valve 60 to the tank. The pressure difference valves 45 and 52 used in the control arrangement according to FIG. 1 are, as has already been pointed out, the same and, as can be seen from FIG. 3, are designed as built-in cartridges. They have a cartridge housing 70, through which a stepped valve bore 71 passes axially. From one end, an adjusting screw 72 is screwed into the valve bore 71, through which the valve bore 71 is closed and which serves to support the control spring 49. This control spring is located in the section of the valve bore 71 with the larger diameter into which the adjusting screw 72 is also screwed. The control spring 49 is supported with its end facing away from the adjusting screw 72 on the piston slide 48, which is guided in the valve bore 71 in an axially movable manner. The free space in the valve bore between the adjusting screw 72 and the piston slide 48 can be referred to as the spring space 75. A star of radial bores 76, which form the outlet 47 of the pressure differential valve, opens into this. At an axial distance from the radial bores 76 and after installation in a block fluidly separated from the radial bores 76 by a sealing arrangement 77, further radial bores 78 pass through the cartridge housing 70 and form the inlet of the pressure difference valve. Along the outside of the cartridge housing 70, even after installation in a block, there is a free fluid connection between the radial bores 78 and the end face 79 of the cartridge housing 70, at which the portion of the valve bore 71 with a smaller diameter emerges to the outside.

The spool 48 is axially guided in the latter section of the valve bore 71 and has an annular groove 80 on the outside through which an annular space is created between it and the wall of the valve bore 71. From the end facing the adjusting screw 72, an axial blind bore 81 is made in the piston slide 48, which extends into the area of the annular groove 80 and is connected there via individual radial bores 82 to the annular groove 80. Further radial bores 83 ensure an open fluidic connection between the bore 81 and the spring chamber 75 and thus also the outlet 47 then when the piston slide 48 abuts with its one end face against a stop of the adjusting screw 72. The piston slide 48 has an outer shoulder 84 with which it can be pressed by the control spring 49 against the inner shoulder of the valve bore 71. When the piston slide 48 rests on the inner shoulder, the annular groove 80 is located between the bore star of the radial bores 78 and the end face 79 of the cartridge housing 70. There is no opening cross section between the radial bores 78 and the annular groove 80. The piston slide is on both sides of the annular groove 80 48 sealingly guided in the valve bore 71 so that the radial bores 78 are fluidly separated from the spring chamber 75 and the annular groove 80 from the space in front of the end face 79 of the valve housing. There is therefore no fluidic connection between the inlet 46 and the outlet 47 of the valve. In operation, the spool 48 is acted upon by the inlet pressure from the end face 79 of the valve housing 70. The compression spring 49 counteracts this and, on an area of the same size as the inlet pressure, the outlet pressure present in outlet 47. There is equilibrium on the spool 48 when the outlet pressure is less than the inlet pressure by a pressure difference equivalent to the force of the compression spring 49. By turning the adjusting screw 72, the pretension of the compression spring 49 and thus the pressure difference between the inlet pressure and the outlet pressure can be changed.

Claims

claims

1. Control arrangement for supplying pressure medium to at least two hydraulic consumers (14, 15, 16) with a demand-flow-controlled (load-sensing-regulated) variable pump (10), the setting of which depends on the highest load pressure of the actuated hydraulic consumers (14, 15, 16) a pump controller (11) can be changed, with two adjustable metering orifices (17, 18, 19), a first of which between an inlet line (13) coming from the variable pump (10) and a first hydraulic consumer (14, 15, 16) and the second is arranged between the feed line (13) and a second hydraulic consumer (14, 15, 16), and with two pressure compensators (23, 24, 25), a first of which is the first metering orifice (17, 18, 19) and the second of the second metering orifice (17, 18, 19) is connected downstream and its control piston can be acted upon on the front side by the pressure after the respective metering orifice (17, 18, 19) in the opening direction, characterized in that d ate the control pistons of the pressure compensators (23, 24, 25) in the closing direction can be acted upon by a control pressure in a rear control chamber (26) which is derived from the supply pressure in the supply line (13) by means of a valve device (45, 52) and changes with the inlet pressure.

2. Control arrangement according to claim 1, characterized in that the difference between the inlet pressure and the control pressure when the variable displacement pump (10) is not yet adjusted to the stop (case of saturation) is not greater than between the inlet pressure and the highest load pressure.

3. Control arrangement according to claim 1 or 2, characterized in that the valve device is a pressure differential valve (45, 52) which with an inlet (46) to the inlet line (13) and with an outlet (47) to the rear Control room (26) of a pressure compensator (23, 24, 25) is connected.

4. Control arrangement according to claim 3, characterized in that the pressure difference valve (45, 52) is set to a fixed pressure difference and has a movable valve member (48) in the sense of opening the fluidic connection between the inlet line (13) and the control chamber (26) on the pressure compensator (23, 24, 25) is acted upon by the inlet pressure and in the sense of closing this connection by the control pressure and by a spring (49).

5. Control arrangement according to a preceding claim, characterized in that the rear control chambers (26) of a plurality of pressure compensators (23, 24) are directly connected to one another, so that the same control pressure prevails in the rear control chambers (26) of these pressure compensators (23, 24).

6. Control arrangement according to one of the preceding claims, characterized by a load signal line (37) into which the highest load pressure of the respectively actuated hydraulic consumers (14, 15, 16) is given via selection valves (36) and by a valve (51), which is a Fluidic connection from the load signal line (37) to the rear control chamber (26) of at least one pressure compensator (23, 24) opens when the difference between the inlet pressure and the highest load pressure falls below a certain value.

7. Control arrangement according to claim 6, characterized in that the valve between the load signal line (37) and the rear control chamber (26) is a to this control chamber (26) opening check valve (51).

8. Control arrangement according to a preceding claim, characterized in that a control pressure for the rear control chamber (26) of the first pressure compensator from the pump pressure by a first valve device (45) (23, 24) and through a second valve device (52) from the pump pressure a control pressure for the rear control chamber (26) of another pressure compensator (25) is derived and that a priority valve (55) is present, via which to maintain a desired pressure difference over the upstream of the other

5 ren pressure compensator (25) arranged metering orifice (19) and thus an adequate pressure medium supply to the corresponding, privileged hydraulic consumer (16) when the displacement pump (10) does not correspond to the required flow rate (case of undersaturation), the control pressure in the rear control chamber (26) first pressure compensator (23, 24) can be raised above the control pressure in the event of io saturation.

9. Control arrangement according to claim 8, characterized in that the priority valve (55) with the inlet line (13) connected to the first connection (56) and one with the rear control chamber (26) of the first pressure compensator (23,

I5 24) connected second connection (57) and has a valve member which in the direction of opening the connection between the first connection (56) and the second connection (57) from the metering orifice in a line section downstream of the preferred hydraulic consumer (16) (19) prevailing pressure and an additional force and in the direction of closing the connection 0 between the first connection (56) and the second connection (57) can be acted upon by the inlet pressure.

10. Differential pressure valve, in particular for use in a control arrangement according to one of the preceding claims, characterized by the following features: a) a valve housing (70) has a valve bore (71), into which an input channel (46) is radial and an axial one At a distance from it an outlet channel (47) opens; b) in the valve bore (71), a piston slide (48) is axially displaceable, with an opening cross-section at the inlet channel (46) can be controlled and which is acted upon on a first end side by the pressure prevailing in the inlet channel (46) and on its second end side by the pressure prevailing in the outlet channel (47); c) in a spring chamber (75) located between the one end face of the piston slide (48) and a closure (72) of the valve bore (71), a compression spring (49) is placed which acts on the piston slide (48) in the direction of reducing the opening cross section; d) the outlet channel (47) opens freely into the spring chamber (75); e) the piston slide (48) is a hollow piston with bores (78), via which an annular space (80), which is formed between the piston slide (48) and the valve housing (70), and which is a circumferential control edge for controlling the opening cross section at the input channel (46), is fluidly connected to the spring chamber (75), and with two sealing sections each sealingly guided in the valve bore (71), one of which sealing section between the input channel (46) and the spring chamber (75) and the seals another sealing section between the fluid path (80, 82, 81, 83) passing through the piston slide (48) and the first end face of the piston slide (48).

1 1. Pressure differential valve according to claim 10, characterized in that the compression spring (49) is supported on a screw plug (72) screwed into the valve bore (71) and closing the valve bore (71).

12. Differential pressure valve according to claim 10 or 11, characterized in that the valve bore (71) in the area of the spring chamber (75) is larger in diameter than in the area on both sides of the input channel (46).

13. Differential pressure valve according to claim 12, characterized in that the valve housing (70) has a built-in cartridge with one on the first end face of the Piston spool (48) is open valve bore (71) and that the piston spool (48) is designed as a stepped piston, from the portion of which with a larger diameter an inner shoulder in the valve bore (71) can be acted upon in the direction of the open side of the valve bore (71) .