EP1135613B1 - Method and control device for controlling a hydraulic consumer - Google Patents

Abstract

The invention relates to a method and a control device for controlling at least one hydraulic consumer. The control device has a pump. The output of the pump can be adjusted in accordance with the load pressure of a consumer. Control occurs by means of a proportional directional control valve (4) that forms a measuring diaphragm. A pressure balance (5) is associated with the control valve and enables the decrease in pressure to remain constant irrespective of load pressure by means of the measuring diaphragm. According to the invention, a lower load pressure is applied to the pump when the pressure balance is opened fully, whereby the decrease in pressure is attenuated by the measuring diaphragm.

Description

The invention relates to a method for control a consumer according to the preamble of the claim 1 and a control arrangement for controlling a hydraulic consumer according to the preamble of the claim Second

Such a control arrangement is, for example, from WO 95/32364 A1 known. In this known solution a variable displacement pump is regulated in such a way that it Output creates a pressure that is around a certain Differential amount above the highest load pressure of all hydraulic Consumer of the tax arrangement lies. Let it go constant pumps in connection with a 3-way flow control valve or variable displacement pumps with changeable Use stroke volume.

For such load-sensing regulations is the case with variable pumps a load sensing controller is provided, which in Meaning of a reduction in the volume of the variable displacement pump of pump pressure and in the sense of increasing the stroke volume the pump of the highest load pressure and one Compression spring can be acted upon. The difference between that Pump pressure and the highest load pressure corresponds to the Force of this compression spring. With these load-sensing circuits is a changeable orifice plate for every consumer and assigned an upstream or downstream pressure compensator, through which the pressure drop across the orifice plate is kept constant becomes so that the hydraulic consumer flow of pressure medium only from the opening cross-section the orifice plate and not the load pressure of the consumer or depends on the pump pressure. At the orifice plates downstream pressure compensators are in which the pump adjusts to the maximum pressure volume has been and the pressure medium flow is not sufficient to the specified pressure drop across the orifice plates of all Consumers maintain the pressure balances of all controlled hydraulic consumers in the closing direction adjusted so that all pressure medium flows to the individual consumers reduced by the same percentage become. With such LUDV systems (load-independent flow distribution) then move all targeted consumers with a percentage around the same value reduced speed.

In LUDV systems, the flow channels are used for reporting the highest load pressure for pump control and the Pressure springs of the individual pressure compensators designed in such a way that the load pressure is reported to the pump controller in an unadulterated manner becomes.

In some applications, the hydraulic pump delivers one Stand-by pressure, which is, for example, 20 bar and to control some consumers or valve arrangements is needed. The one corresponding to the stand-by print Pressure difference must be assigned to the other consumers Orifices are broken down, so that considerable Energy losses occur.

In contrast, the invention is based on the object a method and a control arrangement for control to create at least one hydraulic consumer, where the energy losses are reduced to a minimum are.

This task is accomplished with regard to the procedure the features of claim 1 and in terms of Control arrangement by the features of claim 2 solved.

While in the known load sensing systems The spring of the pressure compensator is always weak was used in order to open the pressure compensator Not to falsify the hydraulic pump reported load pressure, in contrast, the invention is reduced Load pressure to the pump reported. The stroke volume of the pump depending on this reported (reduced) Load pressure set, so that the pressure loss accordingly above the orifice plate is less than the pressure difference on the pump controller (variable pump). That is, the pressure drop over the orifice plate is compared to the conventional Solutions reduced, so that a corresponding energy saving he follows.

When used to carry out the procedure Control arrangement will reduce this for pump control reported load pressure by interpreting the reaches a control piston of the pressure compensating control spring. This is compared with the state of the Technology much higher spring stiffness or preload designed so that the spring force is about that Pressure corresponds to the one reported to the pump controller Load pressure compared to the actually applied load pressure should be reduced. That is, the control arrangement differs essentially compared to conventional solutions by the choice of the spring, so that already existing Control arrangements easily converted can be.

When using a control spring with increased spring stiffness or increased preload, the effective spring force preferably set such that they approximately half of the pressure difference corresponding to Pump controller is present or the state of the art the orifice plate falls off.

The response behavior of the control arrangement is special well if the spring force of the spring over the entire Hub, d. H. from a position with fully closed Control piston up to the fully open position remains about constant. This is particularly easy by means of a suitable pressure medium flow guide, where the resulting from the pressure medium flow Flow forces both in the closing direction and act in the opening direction of the pressure compensator and designed in this way are that they become one with the control spring force Add constants independent of the control piston stroke.

Such a pressure medium flow guide is, for example from the subsequently published P 198 36 564.0 known, the disclosure of which is expressly referred to here is taken.

In the event that the load pressure in the pump Load pressure reporting line via a pressure relief valve is limited, the pressure compensator is preferably with a nozzle bore, through which at complete controlled pressure compensator the load pressure in the load pressure channel is fed. Through the nozzle holes in the pressure scales assigned to the individual consumers the loss flows in several fully controlled Pressure balances and with the pressure relief valve open reduced. This also provides a nozzle bore is in contradiction to the previous load sensing systems used constructions because there - as mentioned at the beginning - always an unadulterated load pressure for pump control has been reported. For this reason the hydraulic resistance of the flow channel to the pump controller chosen as low as possible so that the pressure drop and a falsification of the load pressure when fully open Pressure balance is as low as possible.

The pressure in the load pressure reporting line is preferred via another connection hole in the pressure compensator reported in the spring chamber of the control piston, whereby this connection bore a damping nozzle for damping of pressure fluctuations.

The control arrangement according to the invention can with a Variable pump and associated control unit or with one Constant pump with inlet pressure balance (3-way flow control valve) be executed.

Other advantageous developments of the invention are the subject of further subclaims.

The following is a preferred embodiment the invention with reference to schematic drawings.

Figur 1 ein Schaltschema einer erfindungsgemäßen Steueranordnung;

Figur 2 eine Ventilscheibe mit der Steueranordnung aus Figur 1;

Figur 3 eine Detaildarstellung einer Ventilanordnung mit einer verstellbaren Meßblende und einer nachgeschalteten Druckwaage und

Figur 4 eine Detaildarstellung der Ventilanordnung aus Figur 3.

Show it:

Figure 1 is a circuit diagram of a control arrangement according to the invention;

FIG. 2 shows a valve disk with the control arrangement from FIG. 1;

Figure 3 is a detailed view of a valve assembly with an adjustable orifice and a downstream pressure compensator and

FIG. 4 shows a detailed illustration of the valve arrangement from FIG. 3.

FIG. 1 shows a circuit diagram of a valve disk 2 a valve block with two working ports A, B, one Tank connection T and a pump connection P. To the two working ports A, B is a consumer, for example a hydraulic motor 116 or a double-acting one Cylinder (not shown) connected. Via the hydraulic Circuit can be one of the working connections Connect A, B to pump port P while the other working connection B, A with the tank connection T is connected.

The valve disc also has a control connection LS, over which the load pressure at the assigned Consumer can be tapped.

The pump, not shown, is a variable displacement pump executed, the delivery rate depending on Load pressure of the consumer is regulated. Such load-sensing circuits are adequate in the prior art known, so that a basic description is dispensed with can be. When controlling several consumers via circuits with that shown in Figure 1 Construction is the highest in terms of consumers applied load pressure to the pump and the delivery rate depending on this highest load pressure set.

In the valve disc 2 is a continuously adjustable Directional control valve 4 with a control direction of the Consumer defining direction and a Speedometer forming orifice plate added. The orifice formed by the directional control valve 4 (Speed section) is connected to a pressure compensator 5, whose control piston 40 in its control position the pressure drop across the orifice independent of the load pressure keeps constant. The output connection of the pressure compensator 5 is hydraulic with the directional part of the directional valve 4 connected, depending on the control one of the Working ports A, B supplied with pressure medium and the another is connected to the tank connection T. In the work lines to the work connections A, B are continuous adjustable, unlockable check valve arrangements 6, 8 switched, the reverse flow in its blocking position prevent from the consumer and in the unlocked Pass position a backflow from the corresponding Working connection A or B towards tank connection T. enable.

The directional control valve 4 is activated via Pilot valves 10, 12, through which the end faces of a Directional valve spool 28 of the directional valve 4 with one Control pressure can be applied to this from his to shift the neutral position shown. The Directional valve slide 28 is by two compression springs 30, 32nd biased to its neutral position. The control piston 40 the pressure compensator 5 is in the closing direction of the force a control spring 44 and the highest load pressure Consumer loads through a load pressure channel 22 is tapped by the consumer. The pressure downstream of the directional valve 4 is via a Control line 38 to the one acting in the opening direction End face of the control piston 40 out.

The pilot valves 10, 12 are continuously adjustable executed so that a pressure of the order between the tank pressure and the pressure at the pump connection P can be applied to the end faces of the directional valve slide 28 is. This control pressure is also used to unlock the Check valve assemblies 6 and 8 used.

In a section of the Load pressure channel 22 is a pressure relief valve 45 provided, via which the load pressure in the load pressure channel 22nd is limited. The pressure relief valve 45 is in Spring closing direction and opening direction acted upon by the highest load pressure of the consumers. at If the maximum pressure is exceeded, control oil becomes tank T dissipated. The load pressure channel 22 is furthermore connected to the tank via a tank throttle 47.

Figure 2 shows a specifically designed valve disc 2, in which the circuit according to FIG. 1 is implemented.

As already mentioned, the valve disc 2 has the two working connections A, B and a pump connection P and the tank connection T, which is the valve disc packing of the valve block perpendicular to the plane of the drawing. In addition, the highest load pressure is all over the Valve block controlled consumers to one with the Load pressure channel 22 connected control connection LS performed.

The valve disc 2 has mounting holes for the Directional control valve 4, the directional control valve 28 as Hollow slide is formed. The only in Figure 2 Dashed piston 40 is inside the Directional valve slide 28 out.

Parallel to the directional control valve 4 are 2 in the valve disc the two unlockable check valve arrangements 6, 8 added. Each of the check valve assemblies 6 has a pre-opening main cone 72 which cooperates with an impact piston 92, via which the Main cone 72 for unlocking from its valve seat can be lifted off.

The two pilot valves 10, 12 are in Cartridge design and in the lower in Figure 2 End face of the valve disc 2 screwed in. Both Pilot valves 10, 12 are, for example electrically operated pressure reducing valves, via which the Pressure at the pump connection P to a system pressure at axial outlet connection of the respective pilot valve 10, 12 is reducible. As can be seen from Figure 1, has each pilot valve 10, 12 next to that with the pump connection P connected input port another one with the Tank connection T connected radial connection.

To secure the pressure at the tank connection T is in a check valve 114 intended.

For further details of the check valve assembly 6, 8 and the pilot valves 10, 12 and their Operation is based on the publication DE 196 46 428 A1 referred to the applicant.

DE 196 46 428 A1 discloses a valve arrangement for Actuation of a double-acting consumer at the hydraulic fluid delivered by a pump via a Directional control valve optionally one of two working connections Can be fed, each of which a check valve assembly assigned. Each check valve arrangement is about controlled a pilot valve assigned to this, the Output control pressure also on one end of the Directional valve spool is present. By energizing both Pilot valves can be found in the directional valve spool Clamp the zero position and the two Unlock check valve assemblies so that the Working connections with the check valve arrangements with a tank can be connected and thus a floating position is adjustable.

The structure of the directional valve 4 and the pressure compensator 5 is in the following based on the detailed representation Figure 3 described.

Accordingly, the valve disc 2 for receiving the Directional valve slide 28, a valve bore 50 in which Annular spaces 52, 54, 56, 58 extending radially outward and 60 are formed. As can be seen from FIG. 3, the annular space 52 is dashed over a in Figure 2 indicated load pressure signaling line 62 with the Load pressure channel 22 connected to the control connection LS is led.

The two annular spaces 54 and 56 are over Working channels 66, 68 to the working connections A and B guided.

The annular space 58 is on the one hand Pump line 70 connected to the pump port P and on the other hand via a connecting channel 74 to one Radial connection of the pilot valves 10, 12 out. The Annulus 60 is also hydraulic with the pump connection P connected.

The axial shown in Figure 2 Output connections of the pilot valves 10, 12 are over Control channels 76, 78 with the spring chambers 80, 82 des Directional control valve 4 connected. From there are the control channels 76, 78 further to check valve arrangements 6 and 8 respectively guided.

The directional valve slide 28 is 30.32 in via compression springs biased its basic position shown in Figure 3.

The two compression springs 30, 32 are on locking screws 90 supported, the valve bore 50 in the axial direction to lock.

As already mentioned at the beginning, the Directional valve slide 28 is designed as a hollow piston and has an axial bore 94, which is shown in the illustration Figure 3 from the left end portion of the directional valve spool 28 extends into the area of the ring channel 60. In this axial bore 94 opens a measuring aperture bore star 96, whose breakthroughs as radial bores are formed in the directional valve spool jacket. In the The basic position shown is the orifice hole star 96 arranged between the two annular spaces 58, 60.

In axial distance from the orifice bore star 96 is a directional bore star 98 formed in the shown basic position between the two annular spaces 54, 56 is arranged.

Due to the geometry described above, this can be done Pressure medium from the pump connection P depending on the control of the Directional control valve 28 via the pump line 70, the Orifice hole star 96, the axial bore 94, the Directional bore star 98 and the working channel 66 for Consumer A or accordingly via the annular space 60, the Orifice hole star 96, the axial bore 94, the Directional bore star 98 and the working channel 68 for Consumer B are led.

The control piston 40 is inside the axial bore 94 slidably guided and via the control spring 44 with a Ring face 100 against a stop shoulder Axial bore 94 biased into its closed position. On Control piston spring chamber 104 is via a connecting bore 106, a damping throttle 108 and an opening 109 connected in the directional valve slide 28 to the annular space 52, so that the load pressure the control piston 40 also in Closing direction applied.

According to the enlarged representation in Figure 4, the Control piston 40 an axial blind hole 110 which in the right (Figure 4) end face of the control piston 40 opens. The jacket of the control piston 40 is in the mouth region penetrated radially crown-shaped by openings 112. In A compensation bore star 114 is at a distance from these trained in the position shown in the area of the annular space 54 is arranged.

In the area between the openings 112 and the Bore star 114 is the control piston 40 radial downgraded so that a 112 in the area of the openings first control edge 116 and in the area of Compensation bore star 110 a second control edge 118 is formed.

The radially stepped section 115 is in the Area of the first control edge 116 as an inclined surface trained while in the area of the second Control edge 118 is designed as a radial stage.

The directional valve slide 28 has two according to Figure 4 axially spaced annular grooves 120, 122 which in the Inner peripheral surface of the axial bore 94 are formed.

The two ring grooves 120, 122 are through one Center web 124 separated from each other, the one with the second Control edge 118 cooperates. The one in the illustration 4 right circumferential edge of the first annular groove 120 interacts with the first control edge 116, so that at Axial displacement of the control piston 40 through cooperation the first control edge 116 and the first annular groove 120 a control cross section is opened while by interaction of the second control edge 118 and the Mittelsteg 124 a compensation control cross section is opened. In the basic position shown are the closed both tax cross sections. Regarding further details of this compensation flow guidance on the subsequently published application 198 36 564.0 directed.

The directional bore star 98 opens into the first Ring groove 120. According to FIG. 4, the connecting bore 106 designed as an angle bore, with a radial bore leg 126 of the connection bore 106 in the opening 109 of the directional control valve 28 opens. The radial bore leg 126 is arranged so that the connection between the ring channel 52 and the spring chamber 104 over the entire Stroke of the control piston 40 is open. That is, the Control piston 40 is always in the closing direction by the Force of the control spring 44 and the load pressure in the load pressure channel 22 acted upon.

The axial bore 110 of the control piston 40 opens its left end portion in a nozzle bore 128, the in turn in a radial bore 130 of the control piston 40 transforms.

In the basic position of the control piston shown 40 is the radial bore 130 through the inner peripheral wall of the directional control valve 28 closed. At a Axial displacement of the control piston 40 relative to the directional valve spool 28, the radial bore 130 through a Control edge 132 opened by the breakthrough 109 is formed in the directional valve spool jacket. That is, when the control piston 40 moves against the force of the control spring 44 and the control pressure prevailing in the spring chamber the pressure downstream of the orifice plate into the load pressure channel 22 and thus also reported in the spring chamber 104. The diameter of the nozzle bore 128 is essential smaller than the corresponding connection cross section at the conventional pressure compensators mentioned at the beginning. Latter were always designed so that this connection hole there is only a negligible drop in pressure, so that the load pressure when fully open Control piston 46 is reported to the control pump in an unadulterated manner becomes.

Due to the low trained according to the invention The opening cross section of the nozzle bore 128 is that when the opening is open Pressure relief valve 45 from the load pressure channel 22 to the tank T flowing control oil quantity reduced so that the responsiveness and efficiency of the hydraulic Control can be improved.

The control spring 44 is in the embodiment shown with a large preload or a high spring stiffness executed so that about half the stand-by pressure the variable pump must be applied to the Shift control piston 40 against the force of the control spring 44. That is, at a stand-by pressure of about 20 bar, which in the circuit shown for actuating the Pilot valves 10, 12 is required, the control spring 44th designed as a 10 bar spring. The pressure medium flow along the control piston 44 is by suitable geometric Design of the control cross section described at the beginning and in the opposite direction effective control cross-section designed so that the resultant from the force of the control spring 44 and the flow forces acting on the control piston is a constant that is independent of the control piston stroke. With in other words, the control spring 44 and the flow forces are coordinated so that the result is a horizontal spring characteristic, where the spring force is independent of the stroke of the control piston 40 is.

For a better understanding, the function of the invention Control arrangement explained. First, let's assume that only a single consumer about the Valve block to be supplied with pressure medium. To the pilot valves 10, 12 are activated accordingly, so that a control pressure difference on the end faces of the Directional valve slide 28 acts. According to this control pressure difference the directional valve slide 28 from his spring-biased basic position moved out so that the Orifice hole star 96, for example by a Control edge 133 is opened. Accordingly by the axial displacement of the directional valve slide 28 also opened the directional bore star 98, so that the working channel 66 is supplied with pressure medium from the pump and the working channel 68 with the tank connection T is connected.

The pressure medium passes through the open orifice hole star 96 in the axial bore 94, see above that the control piston 40 of the pressure compensator against the force of the Control spring 44 is acted upon in the opening direction. By the pressure build-up at the inlet of the pressure compensator 5 becomes the control piston 40 brought into its left end position in FIG. 3, so that the compensation cross section and the control cross section are fully controlled on. Doing so the radial bore 130 is opened via the control edge 132, so that the pressure at the input of the pressure compensator 5 over the axial blind bore 110, the nozzle bore 128, the Radial bore 130 and the opening 109 in the load pressure reporting line 62 and thus into the load pressure channel 22 is reported. This reported in the load pressure channel 22 However, load pressure corresponds to the force of the control spring 44 lower than that at the inlet of the pressure compensator and load pressure prevailing in working channel 66. The variable pump then becomes lower depending on this Load pressure controlled.

For the sake of clarity, this is again one Numerical example explained. It is assumed that the Stand-by pressure of the variable pump is 20 bar. The Load pressure at the inlet of the pressure compensator, d. H. in the working channel 66, for example, should be 200 bar. The rule spring 44 is a so-called 10 bar spring (regardless of Stroke). That is, with the fully open Pressure compensator is then a pressure in the load pressure channel 22 from 200 bar - 10 bar = 190 bar reported. The pump will then adjusted accordingly to 210 bar, so that the Pressure drop across the orifice plate is only 10 bar.

In the conventional systems, due to the weak control spring 44 reported a load pressure to the control pump be the pressure at the inlet of the pressure compensator corresponds, d. H. the pump would be in the prior art adjusted to 220 bar so that the pressure drop across the Orifice plate would be 20 bar. By the invention Design of the pressure compensator can thus be considerable Saving energy because the pressure drop over the orifice plate with the pressure compensator fully open 5 is reduced.

In the event that a second via the valve block hydraulic consumer is operated, its load pressure larger than that of the consumer described above is the one assigned to the second consumer Pressure compensator fully opened and this load pressure reported in the load pressure reporting channel and the control pump controlled accordingly. By in the closing direction acting higher pressure is the control piston 40 of shifted the previously described consumer into its control position, in which the control edge 132 the radial bore 130 has contributed. In this control position is at the entrance the pressure compensator the higher load pressure of the second consumer on. This pressure is throttled over the pressure compensator, so that at the pressure compensator outlet (working channel 66) lower load pressure of the first consumer is present. D. that is, via the one assigned to the first and the second consumer Orifice plates have the same pressure difference from about 10 bar.

As already mentioned at the beginning, the Providing the nozzle bore 128 of the control oil flow over the Pressure relief valve to the tank due to the increased hydraulic Resistance minimized, so that the losses the system are reduced to a minimum.

Instead of the variable pump described above a constant pump with 3-way flow control valve is also used become.

A method and a control arrangement are disclosed to control at least one hydraulic consumer. The control arrangement has a pump whose output adjustable depending on the load pressure of a consumer is. Its control takes place via a Proportional directional control valve, one Pressure compensator is assigned, via which the pressure drop over the orifice plate is kept constant regardless of the load pressure. According to the invention, when the pressure compensator is fully open a lower load pressure to the pump is reported, so that the pressure drop across the orifice plate is reduced.

Claims (10)

1. A method for driving a consumer through a control arrangement, having a pump whose output is controlled as a function of a control pressure in a control pressure channel (22) in such a way that the pump pressure is above the control pressure by a predetermined pressure differential, and a directional valve (4) forming a measuring orifice and having a directional valve slide (28) in which a control piston (40) of a pressure compensator (5) is slidably mounted, through which the pressure drop across the measuring orifice is kept constant irrespective of the load pressure of a consumer, the pressure downstream of the measuring orifice being applied to the control piston (40) in the opening direction and a control spring (44) and the control pressure being applied to the control piston (40) in the closing direction, and the pressure downstream of the measuring orifice being indicated to the control pressure channel (22) when the pressure compensator (5) is completely opened, characterized in that when the pressure compensator (5) is completely opened a pressure is indicated to the control pressure channel (22) which is substantially lower however lower by less than the predetermined pressure differential than the load pressure of the consumer.
2. A control arrangement for driving at least one hydraulic consumer, having a pump whose pump pressure is adjustable as a function of a control pressure in a control pressure channel (22) in such a way that the pump pressure is higher than the control pressure by a predetermined pressure differential, having a directional valve (4) for driving the consumer, the directional valve (4) forming a measuring orifice and having a directional valve slide (28), in which a control piston (40) of a pressure compensator (5) is slidably mounted, through which the pressure drop across the measuring orifice can be maintained constant irrespective of the load pressure of a consumer, and to which the pressure downstream of the measuring orifice is applied in the opening direction and to which a control spring (44) and the control pressure is applied in the closing direction, wherein the pressure downstream of the measuring orifice may be indicated to the control pressure channel (22) when the pressure compensator (5) is completely opened, characterized in that the spring constant and/or the bias of the control spring (44) are dimensioned such that when the pressure compensator (5) is completely opened a load pressure of the consumer is indicated to the pump which is substantially reduced however reduced by less than the predetermined pressure differential.
3. A control arrangement according to claim 2, characterized in that the spring force of the control spring (44) corresponds to about half the pressure force corresponding to said pressure differential.
4. A control arrangement according to claim 2 or 3, characterized in that the pressure fluid flow is controlled in such a way that the flow forces acting on the control piston (40) in combination with the force of the control spring (44) exert an approximately constant force on the control piston (40) throughout the control piston stroke.
5. A control arrangement according to claim 4, characterized in that the control piston (40) comprises two control edges (116, 118) allowing a control cross section or a compensating cross section to be driven, enabling pressure fluid flows to be caused acting in the opposite direction.
6. A control arrangement according to any one of claims 2 to 5, characterized in that the control piston (40) comprises a nozzle bore (128) through which a connection between a space downstream of the measuring orifice to the control pressure channel (62, 22) may be opened up, in which a pressure limiting valve (45) is disposed for limiting the control pressure.
7. A control arrangement according to claim 6, characterized by a communication bore (106), through which the spring chamber (104) of the control piston (40) is connectable with the control pressure channel (62, 22).
8. A control arrangement according to claim 7, characterized in that a damping nozzle (108) is provided in the communication bore (106).
9. A control arrangement according to any one of claims 2 to 8, characterized in that the pressure differential corresponds to about 20 bar and the force of the control spring (44) corresponds to about 10 bar.
10. A control arrangement according to any one of claims 2 to 9, characterized in that the pump is a constant pump having an input pressure compensator, or is a variable pump.

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