WO2002093018A1

WO2002093018A1 - Valve arrangement

Info

Publication number: WO2002093018A1
Application number: PCT/DE2002/000010
Authority: WO
Inventors: Thomas Otto Weickert; Martin Karl Schmitt
Original assignee: Bosch Rexroth Ag
Priority date: 2001-05-15
Filing date: 2002-01-04
Publication date: 2002-11-21
Also published as: DE10123699A1

Abstract

The invention relates to a valve arrangement for controlling the supply of a pressure means to or from a consumer, especially a hydrocylinder or a hydromotor. Said valve arrangement has a pressure balance with a downstream reflux valve in a supply line, and a flow control valve in a discharge line, a control side of the pressure balance which is active in the closing direction being connected to the input of the flow control valve by means of a control channel. In the event of a negative load, the pressure means flowing back via the discharge line is fed into the supply line by means of a feeding channel.

Description

description

valve assembly

The invention relates to a valve arrangement for controlling the pressure medium supply and discharge to and from a consumer according to the preamble of patent claim 1.

Valve arrangements of this type are used, for example, to control the working movements of a mobile excavator. In the publication "Oil hydraulics", G. Bauer, G. Teubner, Stuttgart 1998, 7th edition, page 297, a circuit diagram of a fully hydraulic mobile excavator is shown, the hydraulic drive elements of which, for example, a chassis, a slewing gear, a bucket cylinder, etc. can be controlled via a directional control valve, via which the direction of rotation or the retracting or extending movements of the hydraulic working element is predetermined. The pressure medium is supplied via a variable displacement pump, the flow of which is controlled depending on the highest load pressure (LS system). In order to prevent the hydraulic consumer from hurrying ahead in the event of a negative load, for example when the mobile excavator is going downhill, a brake valve is arranged in the outlet of the hydraulic motor of the slewing gear, via which the pressure medium volume flow flowing to the tank is regulated.

The directional control valves of this switching arrangement and the ones described above

Brake valves must be designed for the maximum flow of the hydraulic circuit, so that the valve arrangements require a considerable amount of space due to the pressure medium volume flows which are considerable in mobile hydraulics. As a rule, the valve arrangements are of a relatively complex structure and, due to the large nominal diameters and the additional functions to be integrated, are often very complex, so that a considerable outlay in terms of production technology is required.

To overcome this disadvantage, a valve arrangement is disclosed in EP 02 83 053 B1 in which the control of the hydraulic

Consumer via so-called "Valvistoren" takes place. With such valvistors in principle, it is a 2-way seat valve that can be controlled via a pilot valve.

This pilot valve can, for example, be a continuously adjustable orifice plate, which is followed by a pressure compensator for pressure-independent volume flow control. A cone of the directional seat valve is pressurized to its closed position by a pressure in a rear space. Control slots are formed in the cone, via which a pressure connection of the main stage is connected to the rear space when the cone is lifted off the seat. By actuating the pilot valve, this rear space is connected to the output connection of the valve, so that the plug, depending on the opening of the pilot valve and the pressure at the pressure connection, assumes a control position which determines the pressure medium flow from the pressure connection to the output connection. In other words, with such a valvistor, a large amount of pressure medium can be controlled by a small amount of control oil.

To control hydraulic consumers, a valvistor circuit is proposed in EP 02 83 053 B1, in which two valvistors are assigned to the supply line and the return line. Crosswise control of the valvistors determines the pressure medium supply via a valvistor in one branch and the discharge of the pressure medium from the hydraulic consumer via a valvistor in the other branch, while the other two valvistors in the inlet and outlet lines are not activated. These two valistors are activated when the hydraulic consumer is to be supplied with pressure medium in the opposite direction.

Valvistor control of this type is distinguished from conventional solutions with complex and very large directional control valves by an extremely compact structure, with practically all functions of the directional control valve being able to be carried out by suitable connection of four standard components (valvistors). From the point of view that with such a valvistor control only two valvistors are always in operation and the other two valvistors remain unactuated and due to the fact that the valvistors themselves are very expensive to manufacture, since the main stage has a stepped bore and extremely precise fits - must be carried out, the device engineering effort to implement a valvistor control is also considerable.

Another disadvantage of the known solutions is that with a negative load, for example when driving a mobile excavator downhill at the brake valve or at the valve controlling the flow, a considerable heat load occurs, since these components counteract the pressure of the downhill, for example. 100 bar must work, so that a considerable loss of energy occurs.

In contrast, the object of the invention is to create a valve arrangement which enables hydraulic consumers to be actuated with minimal losses, with minimal expenditure on device technology.

This object is achieved by a valve arrangement with the features of patent claim 1.

According to the invention, the valve arrangement has a flow control valve in a drain line, the pressure at the inlet acting on a pressure compensator arranged in the feed line or a correspondingly acting valve in the closing direction. According to the invention, the drain line can also be connected to the feed line via a feed channel and a feed valve, so that when there is a negative load on the consumer, pressure medium is fed from the drain line into the feed line and cavitation can thus be prevented. Furthermore, due to the high pressure in the drain line at a negative load, the pressure compensator in the feed line is acted on in the closing direction, so that the pressure in the feed line is reduced. In LS systems, this reduced pressure in the supply line is then reported as a load pressure to the variable displacement pump, so that the latter returns to stand-by pressure, for example, so that the pump output is not destroyed, and in thermal energy, as in the conventional solutions must be converted.

In a preferred exemplary embodiment of the invention, a switching valve is assigned to the pressure compensator arranged in the respective feed line. net, via which the control side of the pressure compensator acting in the closing direction can be acted upon either with the pressure at the inlet of the flow control valve in the drain line or with the pump pressure.

When a hydraulic cylinder or a hydraulic motor is activated, the valve arrangement in one exemplary embodiment is designed with two flow control valves and two pressure compensators, the pressure compensator arranged in the discharge line being held in its closed position via the switching valve, while the flow control valve assigned to the supply line is also blocked by suitable activation , so that the control by the flow control valve in the drain line and the pressure compensator in the feed line takes place, the pressure at the input of the flow control valve via the switching valve being applied to the control surface of the pressure compensator which is effective in the closing direction.

The expenditure on device technology can be further reduced if the pressure medium supply to the consumer is controlled via a single metering pressure balance with a downstream directional valve. The directional control valve can be brought from a blocking position into a position in which the output of the single metering pressure compensator is connected to a consumer connection, while the output of the metering pressure compensator is connected to the further consumer connection in the other switching position of the directional control valve. As in the exemplary embodiment described above, the sequence control takes place via the flow control valve in the respectively effective discharge line.

In a structural design of the solution described above, the metering pressure compensator is integrated in a valve slide with a downstream directional valve. The valve arrangement is particularly compact and easy to manufacture if the flow control valves with orifice plate and pressure compensator and the directional control valve with integrated metering pressure compensator are integrated in a common valve housing.

The part of the drain line located downstream of the flow control valve is connected to a tank line. This tank line can be connected via a feed channel with a pressure reducing valve to the pump via which the pressure medium is supplied to the consumer. The Druckreduzierventii and the tank tensioning valves are set so that there is always a minimum tank pressure in the tank line.

When controlling a hydraulic motor, it is preferred if a preload valve is arranged in the feed channel downstream of the pressure-reducing valve, the feed channel being connected to the feed channel in the region between the pressure-reducing valve and the preload valve, so that, in the case of a negative load, in addition to the quantity of pressure medium, which is fed directly from the drain line into the feed line via the feed channel, and additional pressure medium can be fed via the pump into the feed line shut off by the pressure compensator.

In the event that the pressure medium in the drain line is thermally stressed, the drain line can be connected to the connection channel via a thermal valve, which controls the connection when a maximum pressure medium temperature is exceeded, so that the pressure medium supply then only avoids cavitation to avoid it the pump takes place.

According to the invention, it is preferred if the flow control valve is formed by an adjustable measuring orifice and a pressure compensator connected in series. This pressure compensator is preferably acted upon in the closing direction by the pressure at the outlet of the measuring orifice.

In the event that the valve arrangement is to be designed with valvistors, the flow control valve is designed as a pilot valve for a valvistor, via which the sequence control takes place. In this variant, the inflow control is carried out by the pressure compensator arranged in the respective inflow line, which is acted upon by the inlet pressure at the pilot valve in the closing direction. In other words, in contrast to the known solution, only two valvistors are required in such a variant, which is designed with valvistors, while the inflow control takes place via pressure scales which are constructed much more simply or valve devices with a similar effect. The outlay in terms of device technology is thus considerably less than in the valvistor control according to EP 0 283 053 B1. Other advantageous developments of the invention are the subject of the further subclaims.

Two preferred exemplary embodiments of the invention are explained in more detail below with the aid of schematic drawings. Show it:

Figure 1 shows a part of a circuit diagram of a mobile implement with a hydraulic motor and a hydraulic cylinder;

2 shows the circuit diagram for controlling the hydraulic cylinder from FIG. 1;

Figure 3 shows the circuit diagram for controlling the hydraulic motor of Figure 1;

Figure 4 shows an embodiment with valvistors

FIG. 5 shows a schematic diagram to explain the function of a valvator;

FIG. 6 shows a circuit diagram of a simplified valve arrangement according to the invention;

FIG. 7 shows a section through a valve housing into which the valve arrangement from FIG. 6 is integrated and

8 shows a partial illustration of the valve housing from FIG. 7.

FIG. 1 shows a circuit diagram for the control of hydraulic drive elements of a mobile working device, for example an excavator, with the control of a hydraulic motor of a travel drive and a hydraulic cylinder, for example the boom cylinder, being described by way of example.

The circuit shown has a hydraulic motor 2 and a hydraulic cylinder 4, which are supplied with pressure medium via a variable displacement pump 6. The

System is designed as a so-called single-circuit system, with all hydraulic Consumers are supplied by the common variable displacement pump 6. The variable pump 6 is controlled as a function of the highest load pressure of the connected consumers, which is tapped via an LS channel 8. This is connected via an LS pressure relief valve 10 and an LS relief 12 to the tank L, from which the variable pump 6 sucks in pressure medium.

The hydraulic motor 2 and the hydraulic cylinder 4 are controlled via valve arrangements 14 and 16, the elements of which are explained in more detail below with reference to FIGS. 2 and 3. In the exemplary embodiment shown, the valve arrangements 14 and 16 are each combined to form a valve block, each of which has a pressure connection P, a load signaling connection LS, a tank connection T and two working connections A and B.

The pump connection P of the valve arrangement 14 or 16 is connected to the pressure connection of the variable displacement pump 6 via a pressure line 18. A feed channel 21 branches off from the pressure line 18 and is connected to a feed connection S of the valve arrangement 14. A pressure reducing valve 22 is arranged in the feed channel 21, via which the pump pressure is reduced to a feed pressure of, for example, 6 bar. The feed channel 21 is connected via a prestress valve 24 to a tank line 26, via which the tank L is connected to the tank connection T. The tank line .26 has two parallel lines 28, 30, in which tank tensioning valves 32 and 34 are arranged. A cooler 37 for cooling the pressure medium flowing back to the tank L is arranged in the parallel line 28 downstream of the tank tensioning valve 32. The upstream of the cooler 36 tank tension valve 32 opens at a lower pressure than the parallel tank tension valve 34, so that depending on the pressure prevailing in the tank line 26, the entire amount of oil is passed through the cooler 37 or, at higher pressures, through the controlled parallel line 30 directly is returned to tank L.

FIG. 2 shows the valve arrangement 16 for actuating the hydraulic cylinder 4 in an enlarged representation. Accordingly, the valve assembly 16 has two flow control valves 36, 38 connected in parallel, as shown are indicated by dashed lines according to FIG. In the exemplary embodiment shown, each of the flow control valves 36, 38 has a proportionally adjustable measuring orifice 40 or 42, which in the exemplary embodiment shown is formed by an electrically operated proportional valve. The orifice plate 40, 42 is preceded by a pressure compensator 44 and 46, respectively. According to FIG. 2, the flow control valves 36, 38 are each arranged in channels via which the working connection B or A can be connected to the tank connection T. Depending on the control of the control arrangement, one of these channels serves as the inlet line, while the other channel is the outlet line. For the following explanations it is assumed that the pressure medium should flow back from the consumer to the tank via the working connection B and should flow to the consumer via the working connection A. Accordingly, the flow control valve 36 (on the right in FIG. 2) is then arranged in a drain line 48, via which the work connection B is connected to the inlet of the pressure compensator 44. The working connection B of the valve arrangement 16 is connected to an annular space 52 of the hydraulic cylinder 4 via a return channel 50.

Correspondingly, the input of the pressure compensator 46 of the flow control valve 38 is connected to the working connection A via an inlet line 54 and this is connected to a cylinder chamber 58 of the hydraulic cylinder 4 via a pressure channel 56.

According to FIG. 2, the flow control valves 36, 38 are arranged in branch channels 60 and 62, respectively, via which the outlet channel 48 and the inlet channel 54 are connected to a common tank channel 64, which is connected to the tank line 26 via the tank connection T. At the input of the pressure compensator 44 or 46, the pressure in the discharge line 48 or the inlet line 54 is therefore present. The pressure compensators 44, 46 are acted upon in the closing direction by the pressure at the input of the downstream measuring orifice 40 or 42 and in the opening direction by the pressure at the outlet of the respective measuring orifice.

The pressure at the input of the pressure compensator 44, 46 is in each case via a

Control channel 66 and 68 tapped and led to an input port of a switching valve 70 and 72, which is spring-biased in the illustrated

Basic position is blocked. Another input connection of the switching valve tile 70, 72 is connected to the pump connection P via a pump channel 74 connected to the pump connection P.

The drain line 48 and the feed line 54 likewise branch off from this pump channel 74, a metering pressure compensator 76 and 78 being arranged in each case in the branch area. The control surface of the metering pressure compensator 76, 78 effective in the closing direction is acted upon by the pressure at the outlet connection of the switching valve 70 or 72. The control side of the pressure compensator 76, 78 effective in the opening direction is connected to the tank channel 64 via a tank control channel 80 and 82, respectively. In the basic position of the switching valves 70, 72 shown, the pressure compensators 76, 78 are acted upon in the closing direction via the pump channel 74 by the outlet pressure of the variable displacement pump 6. The springs acting in the opening direction are designed so that the pressure compensators 76 and 78 are brought into their blocking position when the hydraulic pump 1 is at standby pressure.

Downstream of each pressure compensator 76, 78 there is an inlet check valve 85 or 87, via which backflow of the pressure medium from the working connection A, B to the pressure compensator 76 or 78 is prevented.

The pressure upstream of the inlet check valves 85, 87 is reported to the LS connection via a signaling channel 89 or 91 and LS check valves 92 or 94.

In parallel to the feed channels 84 and 86, the inlet or outlet line 54, 48 is connected to the tank line via pressure limiting valves 96 and 98, so that in the event of a pressure increase in the inlet line 54 or in the outlet line 48 pressure medium to the tank L can be released ,

To extend the cylinder 4, the switching valve 72 is actuated and the orifice 40 is opened. When the switching valve 72 is actuated, the pressure at the inlet of the pressure compensator 44 is applied to the control surface of the pressure compensator 78, which pressure corresponds to the tank pressure specified by the tank tensioning valves 32, 34. The pressure compensator 78 opens that

Pressure compensator 76 and the orifice plate 42 are closed. That of the Control pump 6 pumped pressure medium is fed via the pressure port P, the pressure compensator 78 and the check valve 86 to the working port A and from there via the pressure channel 56 into the cylinder chamber 58. The pressure medium displaced from the annular space 52 flows via the working connection B, the discharge line 48, the pressure compensator 44 and the opened measuring orifice 40 into the tank channel 64 and from there via the tank connection T into the tank line 26. The pressure compensator 44 is in a control position a, in which the pressure drop across the orifice 40 is kept constant, regardless of the pressure in the branch channel 60 (drain line 48).

According to FIG. 2, the tank channel 64 is connected to the drain line 48 and the feed line 54 via feed channels 84, 86. In each of the feed channels 84, 86, a feed valve 88 or 90 designed as a check valve is arranged.

With a pulling load (upward in FIG. 2), there is a substantially higher pressure in the discharge line 48 than in the "regular" operation of the hydraulic cylinder 4, while the pressure in the supply line 56 is low, so that pressure medium has to flow into the cylinder chamber 58 , The pressure of, for example, 100 bar present in the discharge line 48 is then also present at the inlet of the pressure compensator 44. This pressure is fed via the control channel 66 and the switched switching valve 72 to the control surface of the metering pressure compensator 78 which is effective in the closing direction, so that it is adjusted against the force of its compression spring and the pressure in the feed line 54 is throttled. This reduced pressure is reported via the signaling channel 91 and the LS check valve 92 and the LS connection in the LS channel 8. Due to this reduced load pressure, the variable displacement pump 6 is then retracted. This retraction of the variable pump saves a considerable amount of energy compared to conventional solutions, since in these the variable pump continues to deliver at the preset pressure.

The pressure medium then flows through the flow control valve 36 with a constant volume flow into the tank channel 64. Since the pressure in the inlet line 54 is substantially lower than in the outlet line 48, the feed valve 90 opens, so that the pressure medium is fed directly into the inlet line 54 can. Cavitation in the cylinder space 58 can thus be avoided. The system is thus operated with almost 100% regeneration, so that this solution is energetically superior to conventional solutions.

FIG. 3 shows an enlarged illustration of the valve arrangement 40 for controlling the hydraulic motor 2. This valve arrangement 40 corresponds in FIG

Basic structure of the valve arrangement 16 described with reference to FIG

Control of the hydraulic cylinder 4, so that only the additional in the following

Components are described.

As already mentioned at the beginning, the connecting channel 20 is connected to the

Supply connection S connected. This is connected via a channel 102 to the feed channels 86, 88. In contrast to the exemplary embodiment described above, these are not connected directly to the tank channel 64. This connection is made via a thermal valve 104 which, in its spring-loaded basic position, connects the tank channel 64 and thus the return line 48 (54) to the channel 102. When a strongly heated pressure medium flows through it, the thermal valve 104 closes the connection between the tank channel 64 and the channel 102, so that pressure medium can be fed into one of the feed channels 86, 88 only via the connecting channel 20 and not via the tank channel 64 ,

According to FIG. 3, a pressure-limiting channel 106, 108 branches off from the drain line 48 and from the feed line 54, which is connected to the tank control channel 80 and 82, respectively. A pressure relief valve 110 or 112 is arranged in each pressure relief passage 106, 108 and, in its spring-loaded basic position, shuts off the associated pressure relief passage 106 or 108. The pressure in the pressure-limiting channel 106 or 108 is reported to a control surface which acts on the switching valve 70 or 72 in the basic position. Furthermore, a throttle 114 or 116 is provided in each pressure limiting channel 106, 108.

To control the hydraulic motor, ie to apply pressure via the working port A, the orifice plate 40 is opened and the switching valve 72 is switched over, while the orifice plate 42 and the further switching valve 70 remain in their closed positions. The pressure medium is supplied via the pressure connection P, the metering pressure compensator 78 and the inlet check valve 87 and promoted the working connection A to the hydraulic motor 2. The returning pressure medium flows via the working connection B, the pressure compensator 44 'of the flow control valve 36 and the downstream controlled measuring orifice 40 to the tank channel 64 and from there via the tank connection T back into the tank line 26.

With a negative load, a high pressure (for example 100 bar) builds up in the discharge line 48, which is reported via the control channel 66 and the switched switching valve 72 to the control surface of the metering pressure compensator 78 - this is controlled and the variable displacement pump 6 is retracted. The pressure medium in the discharge line 48 is introduced into the tank channel 64 via the pressure compensator 44 and the activated measuring orifice 40 with a volume flow that is essentially independent of the load pressure. If this pressure medium has a permissible working temperature, the thermal valve 104 is opened so that the pressure medium is fed into the channel 102 via the thermal valve 104. In this channel 102 there is at least the minimum tank pressure set via the pressure reducing valve 22 and the tank tensioning valves 32, 34. The pressure medium can then be fed into the feed line 54 via the feed channel 86 and the open feed valve 90, with sufficient pressure medium being able to be fed into the connecting channel 20 with stand-by pressure by means of the additional pressure medium supply via the variable displacement pump 6 in order to prevent any cavitation , When the pressure medium is very hot, the thermal valve 104 closes, so that the pressure medium flowing back via the discharge line is returned to the tank L via the tank channel 64 and the tank line 26. The cavitation is prevented by the pressure medium supply via the variable pump 6.

In the above-described solutions according to the invention, the hydraulic consumers are actuated via a sequence control, and when a pressure exceeding a predetermined pressure level occurs in the sequence, the variable displacement pump 6 can be retracted and pressure medium can be fed into the inlet, so that an almost 100% regeneration is possible. FIG. 4 shows a variant of the circuits shown in FIGS. 2 and 3. In this variant, the flow control valves 36 are designed in the respective sequence as a pilot valve of a valvistor 114 or 116.

The basic principle of a valvistor is shown in FIG. 5. As mentioned at the beginning, the main stage 118 is designed as a 2/2-way seat valve, the cone 120 of which is seated on a valve seat 122. A connection between an input connection P and an output connection T can be controlled via the stepped cone 120.

The end face of the cone 120 which is effective in the closing direction is arranged in a rear space 124 which is connected to the output connection T via a control line. A pilot valve 128 is arranged in the control line 126 and forms an adjustable measuring orifice.

The cone 120 is provided with a longitudinal slot 130 which is blocked in the closed position of the cone 120 by a control edge 132 in the annular space, so that there is no direct connection between the input port P and the rear space 124. When the pilot valve 128 is opened to a predetermined orifice cross section, control oil flows out of the rear space 124. The cone 120 lifts off the valve seat 122 and the control edge 132 opens the longitudinal slot 130, so that pressure medium can enter the rear chamber 124 from the inlet port P. The control cone 120 assumes a control position in which the control oil supply via the control slot 130 is equal to the control oil discharge from the rear space 124 via the pilot valve 128. In other words, with such a valvistor, large pressure medium volume flows can be controlled with very small amounts of control oil.

For example, a flow control valve 36, 38 can be used as the pilot valve 128 according to FIG. 5, as was described in the exemplary embodiments described above.

FIG. 4 shows the circuit symbol of the main stage 118, as shown in FIG

Figure 5 is explained. The adjustable measuring orifices mechanically connected to each other in the circuit symbol of the main stage 118 stand for the opening cross-section controlled by the cone 120 (measuring orifice flowing upward of the pressure compensator of the main stage) and for the cross-section of the longitudinal slot 130, which is controlled via the control edge 132. When using valvistors, the input of the flow control valve 36 - in the present case the input of the pressure compensator 44, is connected to a control line 134 via which the pressure compensator Main stage 118 of valve 114 is acted upon by a pressure effective in the closing direction.

Accordingly, the two main stages 118 of the two valvistors 114, 116 are connected with their input connection P to the drain line 48 and the feed line 54 and the output connections T are each connected to the common tank channel 64.

The control channel 66 leading to an input port P of the switching valve 72 is also connected to the control line 134.

The pressure compensator of the main stage 118 of the valvistor 116 is accordingly acted upon in the closing direction via a control line 136, the pressure of which is also present at the input of the pressure compensator 46 and via the control channel 68 also at an input of the switching valve 70. The two tank control channels 80 and 82 are connected to the tank channel 64 as in the previously described exemplary embodiments. For the rest, the exemplary embodiment shown in FIG. 4 corresponds to the circuit according to FIG. 2. The circuit according to FIG. 3 can accordingly also be implemented with the two valvistors 114 and 116.

To extend the hydraulic cylinder 4, the switching valve 72 is switched over and the adjustable measuring orifice 40 is opened. The switching valve 70 and the other adjustable measuring orifice 42 remain in their basic positions shown.

The pressure medium flows via the open metering pressure compensator 78, the inlet check valve 87 to the working connection A and from there into the cylinder space 58. The pressure medium displaced from the annular space 52 flows via the working connection B, the drain line 48 to the connection P of the main stage 118 of the valve 114. Bei activated orifice 40, which takes over the function of pilot valve 128 in FIG. 5, is activated in the closing direction. same pressure reduced by the control oil flow in the control line 134 - the main stage 118 opens and assumes a control position in which the pressure medium can flow from the inlet port P to the tank port T into the tank channel 64.

With a pulling load, the pressure in the drain line 48 increases. The pressure in the control line 134 is also increased accordingly, so that this increased pressure acts on the metering pressure compensator 78 in the closed position via the control channel 66. The closing metering pressure compensator 78 reduces the pressure in the feed line 54 and reports a correspondingly reduced pressure via the signaling channel 91 and the LS check valve 92 in the LS channel 8 - the variable displacement pump 6 moves back to stand-by pressure. The control oil volume flow through the flow control valve 36 is kept constant by the combination of the pressure compensator 44 and the open measuring orifice 40, so that the pressure medium is in its control position in the main stage 118 of the valve 114 from the input port P to the tank port T and from there into the tank channel 64 can flow off. Since the pressure in the feed channel 54 is substantially lower than the pressure in the tank channel 64, pressure medium can be fed into the feed line 54 via the feed channel 86 and the feed valve 90, just as in the exemplary embodiments described above.

In the above-described exemplary embodiments, a metering pressure compensator 76, 78 and a switching valve 68, 70 are provided in each branch, which interact cross-wise with the respective flow control valve 36 or 38 in the other branch in order to control the pressure medium inflow and outflow. FIG. 6 shows a simplified exemplary embodiment in which a single, common metering pressure compensator 78 is assigned to the two branches, which is followed by a directional control valve 136.

The directional control valve 136 has three switching positions in the exemplary embodiment shown, the inlet line 54, the outlet line 48 and the pressure line 18, into which the metering pressure compensator 78 is connected, being shut off in the central, spring-preloaded basic position (0). For the supply of pressure medium via the inlet line 54, the directional control valve can be moved into a switching position (b) electrically or by means of a hydraulic control pressure. bring in which the outlet of the pressure compensator 78 is connected to the inlet line 54, while the connection between the pressure line 18 and the return line is blocked. In this switching position, the control channel 66 connected to the input of the flow control valve 36 is connected to a line 138, so that the control surface of the metering pressure compensator 78 effective in the closing direction is acted upon by the pressure at the input of the flow control valve 36 and thus the pressure in the discharge line 48. As in the exemplary embodiments described above, the control surface of the metering orifice 78 which is effective in the opening direction is acted upon by the tank pressure, ie by the pressure in the tank channel 64.

In the other switching position (a) shown, the control surface of the metering orifice 78 which is effective in the closing direction is acted upon by the pressure in the other control channel 68, which is connected to the inlet of the flow control valve 38. In this switching position, the output of the metering pressure compensator 78 with the right branch, i. H. connected to the line designated as "drain line" 48, so that the consumer is supplied with pressure medium via the working connection B and this line acts as an inlet line.

In the exemplary embodiment described above, the pressure medium supply and discharge to and from the consumer is thus controlled via the directional valve 136, the pressure compensator 78 always being acted upon in the closing direction by the pressure in the respectively effective drain line and in the opening direction by the tank pressure.

Similar to the exemplary embodiment shown in FIG. 1, the tank pressure is prestressed to a predetermined value, for example 8 bar, by means of a tank tensioning valve 32. A pressure reducing valve 22 is provided in the feed channel 21 for setting the tank feed pressure.

The other elements of the circuit shown in Figure 6 correspond to the embodiment of Figure 2, so that for the sake of simplicity reference is made to the corresponding statements. For pressure medium supply via the working connection A, the directional valve 136 is brought into the switch position designated (b), so that the pressure medium is guided to the consumer via the working connection A via the pressure compensator 78, the directional valve 136, the inlet line 54 and the inlet check valve 87. The pressure medium flowing back from the consumer is returned to the tank L via the working connection B, the discharge line 48, the pressure compensator 44, the metering orifice 40 and the tank channel 64.

When pressure builds up in the drain line (pulling load), the metering pressure compensator 78 is controlled - similar to the embodiment described with reference to FIG. 2 - by the pressure in the drain line 48 tapped via the control channel 66, so that the pressure in the feed line 54 is throttled. The reduced pressure is then tapped via the message channel 91 and reported in the LS channel 8, so that the variable pump 6 is retracted. The pressure medium flowing back via the working connection B can then be fed directly into the feed line 54 via the feed valve 90 in the manner described at the beginning, so that cavitation can be prevented.

In special applications, the directional control valve 136 - as in FIG

6 indicated - be designed as a proportional valve. In principle, however, it is sufficient if the directional valve 136 is designed as a switching valve with three switching positions.

FIGS. 7 and 8 show the constructive implementation of the circuit according to FIG. 6. In this exemplary embodiment, practically all of the components described above are integrated in a common valve housing 140. The valve housing 140 has a first valve bore 142 in which a valve slide 144 of the directional valve 136 is guided. A piston 146 of the metering pressure compensator 78 is integrated in the valve slide 144.

A further bore 148 is formed in the valve housing 140 parallel to the valve bore 142, in which a valve body 150 is guided, via which the two measuring orifices 40, 42 are formed. Pressure compensator pistons 152, 154 of the pressure compensators 44 and 46 are arranged axially displaceably in the valve body 150. As will be described in more detail below, the two inlet check valves 85, 87 are also integrated into the valve housing 140. The directional control valve 136 and the two measuring orifices 40, 42 are actuated hydraulically via control channels 156, 158, which are each connected to control rooms of the directional control valve 136 and the measuring orifice 44 or 46.

The valve slide 144 is biased by springs 160, 162 and the valve body 150 by means of centering springs 164, 166, the spring spaces of the

Centering spring 164 and spring 162 with the pressure in the control channel 156 and the spring spaces of the centering spring 166 and the spring 160 with the pressure in the control channel 158.

With regard to further details of the valve arrangement shown in FIG. 7, reference is made to the detailed illustration according to FIG. 8.

Accordingly, five annular grooves 168, 170, 172, 174 and 175 are formed in the valve bore 142. The central annular groove 172 is connected to the variable displacement pump 6 via the pressure line 18. The valve spool 144, which is designed as a hollow slide, has a casing bore star 178 which, in the basic position shown, is arranged in the region between the annular grooves 168 and 170. At the axial distance from the casing bore star 178, control bores 180 are formed in the valve slide, which open into the annular groove 172.

In the position shown, a jacket bore 182 is formed in the valve slide 144 between the two annular grooves 174, 176, via which the pressure in one of the annular grooves 174, 176 can be applied to the right end face of the piston 146 in FIG. 8. The other end face of the piston 146 can be acted upon via a radial bore 184 with the pressure in a control line 186 via which the tank pressure is tapped, so that the left end face of the piston 146 in FIG. 8 is acted upon by the tank pressure.

The piston 146 of the metering pressure compensator 78, which is guided in an inner bore 188 of the valve slide 144, has two axially spaced ring collars 180, 190, which delimit an annular space 192 with the inner circumferential wall of the inner bore 188. The axial spacing of the annular collars 189, 190 is selected so that the annular space 192 connects the casing bore star 178 to the control bores 180 in the basic position shown.

The end faces of the valve slide 144 are closed via end pieces 194 and 196. A control spring 198 is supported on the end piece 194, via which the piston 146 is biased against the end piece with an end part 200 arranged to the right of the annular collar 190. The end part 200 is set back radially with respect to the annular collar 190, so that pressure medium can pass through the casing bore 182 and the end face of the end part 200 can be subjected to the pressure in the annular groove 174 or 176. This pressure acts against the force of the control spring 198 and the pressure in the spring chamber of the piston 146.

Four annular grooves 202, 204, 206, 208 are formed in the bore 148 of the valve housing 140 which receives the valve body 150, the annular grooves 202, 208 being connected to the working connections A and B, respectively. The annular groove 202 is connected to the annular groove 168 via the inlet line 54 and the inlet check valve 87 integrated in the valve housing 144, while the annular groove 208 is connected to the annular groove 170 via the outlet line 48 and outlet check valve 85. The higher pressure in the inlet or outlet line is tapped via the LS signaling channel 91.

The bore 148 also has two grooves 210, 212, which over the

Control channels 66 and 68 are connected to the ring grooves 176 and 174, respectively. The pressure in the ring grooves 202 and 208 can be tapped via the grooves 210 and 212 upon a corresponding displacement of the valve body 150 and can be transmitted into the associated ring grooves 176 and 174.

In the jacket of the valve body 150, central tank openings 214 are formed, which open into a circumferential groove of the valve body 150. Through this circumferential groove, an annular gap 216 is formed, which is between the

Ring grooves 204 and 206 extends and which is connected to the tank channel 64, so that the tank pressure is present in the annular gap 216. The geometry of this Annular gap 216 is selected such that the annular grooves 204 and 206 are connected to the tank channel 54 in each axial position of the valve body 150.

The two pressure compensating pistons 152, 154 are guided in the interior of the valve body 150, and these are prestressed via a central, common control spring 218 against two end pieces 220, 222 closing the end of the valve body 150. Four further jacket bore stars 226, 228, 230, 232 are formed in the jacket of the valve body 150, the bore stars 228, 230 and 232, 226 being configured symmetrically to the tank openings 214. The jacket bore stars 226, 228 are assigned to the pressure compensator 46 and the jacket bore stars 230, 232 to the pressure compensator 44 (see FIG. 7). The two pressure compensator pistons 152, 154 each have two axially spaced collars 234, 236 and 238, 240, respectively, through which an annular space 242 and 244 is delimited. In the basic position of the pressure compensator pistons 152, 154 shown on the right in FIG. 8, this annular space extends in such a way that the two casing bore stars 230 and 232 or 226 and 228 are connected to one another. In the illustration according to FIG. 8, the left pressure compensator piston 154 is shown in a control position in which the casing bore star 226 is closed. The pressure in the annular space 242 or 244 can be tapped via inner bores 246 or 248 and can be passed to the outer end face of the pressure compensator piston 152 or 154, which is remote from the control spring 218.

To supply pressure medium via the working port A, a control pressure is applied to the control channel 156, so that the valve slide 144 and the valve body 150 in the illustration according to FIG. 7 are shifted to the left by the control pressure against the force of the springs 160 and 166. As a result of this axial displacement, the annular groove 172 is connected to the annular groove 168, which is now open, via the control bores 180, the annular space 192 and the casing bore star 178. The pressure medium flows via the inlet line 54, the inlet check valve 87, the annular groove 202 to the working connection A. The pressure medium flowing back from the consumer enters the annular groove 208 via the working connection B, which due to the axial displacement of the valve body 150 via the casing bore star 232 with the annular space 242 That is, the casing bore star 232 and the annular groove 208 practically form the measuring orifice 40. The pressure in the annular Room 242, ie at the outlet of the orifice plate 40 is tapped via the inner bore 246, so that the pressure compensator piston 152 (FIG. 7) is acted against the force of the control spring 218. In addition to the control spring 218, the pressure in the tank line 64 acts on the left line surface of the pressure compensator piston 152.

The pressure compensator piston 152 sets itself into a control position in which the pressure drop across the measuring orifice is kept constant, regardless of the pressure in the discharge line 48, i. H. in this control position, the opening cross section of the casing bore star 232 is opened or closed accordingly by the collar 236.

The pressure medium returning from the working connection B then enters through the annular space 242, the casing bore star 232 and the annular groove 206 into the annular gap 216 and can flow out through the tank openings 214 to the tank. The pressure in the drain line 48 is tapped via the groove 212 and the control channel 66 and is present in the annular groove 174. This pressure is guided via the casing bore 182 to the right end face of the piston 146 in FIG. 8, so that the pressure in the discharge line 48 is applied to it in the closing direction. With the above-described pressure increase in the return line 48, the piston 146 is then shifted to the left against the force of the control spring 198, so that the control bores 180 are closed and the pressure in the supply line 54 is throttled. This reduced pressure is reported via the signal line 91 to the variable displacement pump 6, so that the latter is retracted. In this operating state, as described at the beginning, the pressure medium can be fed directly into the feed via the feed valve 90 (FIG. 7).

In all of the above-described exemplary embodiments, a proportional flow controller can also be used instead of the flow control valve 36, 38 with a pressure compensator 44 or 46 with a proportionally adjustable measuring orifice 40, 42.

The feed valves 88 and 90 in the form of check valves and the pressure relief valves 96 and 98 can be integrated into the main valvistor stage 118. Disclosed is a valve arrangement for controlling a pressure medium supply and discharge to and from a consumer, in particular a hydraulic cylinder or a hydraulic motor. The valve arrangement has a pressure balance in a feed line with a downstream check valve and a flow control valve in a discharge line, a control side of the pressure balance effective in the closing direction being connected via a control channel to the input of the flow control valve. The pressure medium flowing back through the drain line can be fed into the feed line via a feed channel if the load is negative.

LIST OF REFERENCE NUMBERS

2 ydromotor

4 hydraulic cylinders

6 variable pump

8 LS channels

10 LS pressure relief valve

12 LS relief

14 valve arrangement (hydraulic motor)

16 valve arrangement (hydraulic cylinder)

18 pressure line

20 connecting channel

21 feed channel

22 pressure reducing valve

24 preload valve

26 tank line

28 parallel line

30 parallel line

32 tank tension valve

34 tank cocking valve

36 flow control valve

37 cooler

38 flow control valve

40 orifice plate

42 orifice plate

44 pressure compensator

46 pressure compensator

48 drain pipe

50 return channel

52 annulus

54 Inlet line

56 pressure channel

58 cylinder chamber

60 branch channel

62 branch channel tank channel

control channel

switching valve

pump channel

Zumeßdruckwaage

Tank control channel

feed channel

Inlet check valve

feed channel

Inlet check valve

feed valve

signaling channel

feed valve

signaling channel

LS-return valve

Pressure relief valve

channel

thermovalve

Pressure relief channel

Pressure relief valve

valvistor

main stage

cone

valve seat

backcourt

control line

pilot valve 130 longitudinal slot

132 control edge

134 control line

136 way valve

138 line

140 valve housing

142 valve bore

144 valve spool

146 pistons

148 hole

150 valve body

152 pressure balance pistons

154 pressure balance pistons

156 control channel

158 control channel

160 feather

162 spring

164 centering spring

166 centering spring

168 ring groove

170 ring groove

172 ring groove

174 ring groove

176 ring groove

178 jackets

180 control holes

182 casing bore

184 radial bore

186 control line

188 inner bore

189 ring collar

190 bundle of rings

192 annulus

194 end piece

196 end piece

198 rule spring 200 end part

202 ring groove

204 ring groove

206 ring groove

208 ring groove

210 groove

212 groove

214 tank breakthroughs

216 annular gap

218 control spring

220 end piece

224 end piece

226 mantle bore star

228 mantle bore star

230 mantle bore star

232 mantle bore star

234 fret

236 fret

238 fret

240 bunch

242 annulus

244 annulus

246 inner bore

248 inner bore

Claims

Expectations

1. Valve arrangement for controlling a pressure medium supply and discharge to or from a consumer, in particular a hydraulic cylinder (4) or a hydraulic motor (2), with a valve device (14, 16) via which one to the consumer (2, 4) leading inlet line (54) can be connected to a pump (6) and a return line (48) to a pressure medium reservoir (L), characterized in that in the inlet line (54) a metering pressure compensator (78) and in the outlet line (48) a flow control valve (36) is arranged, wherein a control surface of the metering pressure compensator (78) which is effective in the closing direction can be acted upon via a control channel (66) with the input of the flow control valve (36) and the pressure medium flowing back via the outlet channel (48) by means of a feed channel (86) can be fed into the feed line (54) with a feed valve (90).

2. Valve arrangement according to claim 1, wherein the metering pressure compensator (78) is assigned a switching valve (72), in the switching positions of which the control surface can be connected to a line carrying the pump pressure or to the control channel (66).

3. Valve arrangement according to claim 1 or 2, wherein the drain line (48) is connected to a tank line (26) in which. at least one tank tensioning valve (32, 34) is arranged.

4. Valve arrangement according to claim 3, wherein the tank line (26) via a feed channel (21) with pressure reducing valve (22) is connected to the pump (6).

5. Valve arrangement according to claim 4, wherein a biasing valve (24) is arranged in the feed channel (21).

6. Valve arrangement according to claim 4 and 5, wherein the feed channel (21) in the region between the biasing valve (24) and pressure reducing valve (22) via a connecting channel (20) is connected to the feed channel (84,86).

7. Valve arrangement according to claim 6, wherein the drain channel (48) via a thermal valve (104) with the connecting channel (20) can be connected.

8. Valve arrangement according to one of the preceding claims, wherein the flow control valve (36,38) is a proportionally adjustable orifice

(40.42) and a pressure compensator (44.46).

9. Valve arrangement according to one of the preceding claims, wherein the flow control valve (36,38) is a pilot valve (128) of a vavistor (114,116) for opening the connection between the drain channel (48) and the tank line (26).

10. Valve arrangement according to one of the preceding claims, wherein each consumer (2,4) for supply and sequence control are assigned two flow control valves (36, 38) and metering pressure compensators (76, 78) arranged in parallel branches, which can be controlled crosswise, so that the pressure medium supply can be controlled via a metering pressure balance (78) in one branch and the pressure medium discharge can be controlled via the flow control valve (36) in the other branch.

11. Valve arrangement according to one of claims 1 to 9, wherein a single metering pressure compensator (78) is provided, in which a directional valve (136) is connected, which in a position (0) shuts off the supply line (54) and the control channel (66) and in another position (a, b) connects the inlet line (54) to the outlet of the metering pressure compensator (78), and the control surface effective in the closing direction is acted upon by the pressure in the control channel (66).

12. Valve arrangement according to claim 11, wherein each consumer (2,4) for supply and sequence control are assigned two flow control valves (36, 38) arranged in parallel branches, to which a metering pressure compensator (78) with a downstream directional valve (136) is assigned, wherein The pressure medium supply to the consumer can be controlled via the directional valve (136) and the pressure medium discharge can be controlled via the flow control valve (36, 38) connected to the control channel 66.

13. Valve arrangement according to claim 12, wherein the metering pressure compensator (78) is integrated in a valve slide (144) of the directional valve (136).

14. Valve arrangement according to claim 8 and 13, wherein the flow control valves (36,38) with orifice plate (40,42) and pressure compensator (44,46) with the

Directional control valve (136) are integrated in a common valve housing (140), measuring orifice (40, 42) and directional control valve (136) can be actuated by the same control pressure.

15. Valve arrangement according to one of the preceding claims, wherein the pump is a variable displacement pump (6) which is assigned to several consumers and is controlled via the highest load pressure of the consumers, which is via an LS channel (8) at the outlet of the metering pressure compensator (76,78 ) is tapped.