WO2016091528A1

WO2016091528A1 - Hydraulic valve arrangement, hydraulic valve block with such a valve arrangement, and hydraulic drive comprising such a valve block

Info

Publication number: WO2016091528A1
Application number: PCT/EP2015/076408
Authority: WO
Inventors: Christoph Keyl; Steffen Rose
Original assignee: Robert Bosch Gmbh
Priority date: 2014-12-08
Filing date: 2015-11-12
Publication date: 2016-06-16
Also published as: DE102015209657A1

Abstract

The invention relates to a hydraulic valve arrangement for supplying a pressure medium to a hydraulic load, having two working connections for connecting to the load. The valve arrangement comprises a first inlet flow measuring diaphragm, via which a first inlet flow can be routed towards the load from the first of the working connections, and a second return flow measuring diaphragm, which is separate from the first inlet flow measuring diaphragm and via which a first return flow can be routed from the load via the second working connection. The valve arrangement also comprises a second inlet flow measuring diaphragm, via which a second inlet flow can be routed towards the load from the second of the working connections, and a first return flow measuring diaphragm, which is separate from the second inlet flow measuring diaphragm and via which a second return flow can be routed from the load via the first working connection. The invention further relates to a hydraulic valve block comprising at least one such valve arrangement and to a hydraulic drive comprising a hydraulic pump, at least one hydraulic load, and at least one such valve arrangement or valve block.

Description

HYDRAULIC VALVE ASSEMBLY, HYDRAULIC VALVE BLOCK WITH SUCH A VALVE ASSEMBLY, AND HYDRAULIC ACTUATOR WITH ONE

SUCH VALVE BLOCK

description

The invention relates to a hydraulic valve assembly according to the preamble of patent claim 1, a hydraulic valve block according to claim 12, and a hydraulic drive according to claim 15.

Generic valve arrangements find particular in the field of mobile

Work machines, such as a hydraulic excavator, a tractor, a backhoe loader or a forklift application. In particular, load-sensing systems with upstream individual pressure compensators - also called LS systems - load-sensing systems with downstream individual pressure compensators - also called LUDV systems - and throttle controls in open-center circuits with a constant or

demand-adjusted volume flow supply prevailed. Conventional valve assemblies often have the property that the supply and return flow of a consumer-defining throttle cross-sections are changed together by only one spool. In order to avoid cavitation in the feed or return line under all circumstances during operation of the mobile working machine, the associated inlet and return flow orifices must be on

Spool be provided with specially matched notch geometries. However, a tuning of the notch geometries of the inlet and return flow orifices is dependent on a kinematics moved by the hydraulic consumer. For example, in the case of a working arm of an excavator, the positions of gravity and articulation points,

Transmission ratio of coupling mechanisms, the applied loads and the geometric dimensions of the consumer an influence on the optimal design of notch geometries. For this reason, the notch geometries of the inlet and return flow orifices must be individually adapted to each consumer. This process is trial-based and iterative and therefore very expensive and costly. In the publication DE 10 2009 058 404 AI therefore a valve arrangement is proposed, which has dissolved control edges for independent control of the inlet and return. Here, the respective control pressures are switched via a logic slide so that an independent positioning of the two control spool (inlet, return) is possible. Thus, an independent adjustment of the inlet and return flow orifices is possible. The individual geometric adjustment of the inlet and return flow orifices can thus be replaced by an individual control of the inlet and return flow orifices, which can also be applied by software technology. Disadvantage of the presented valve arrangement is that the realization of a free circuit is not possible. An application in the tractor is thus obsolete.

Furthermore, valve assemblies are known in dissolved construction, in which the inlet valves are designed via electrically controllable valves in seat design and each consumer port is associated with an inlet valve. The respective return valves are represented by lowering brake valves. These are pushed open by the pressure in the inlet channel and it is set in a return flow rate so that the pressure in the inlet channel is set or regulated to a value in the broadest sense of the biasing force of the return spring and a dependent on the load pressure in the return channel intermediate pressure in the spring back of the lowering brake valve spool. A disadvantage of valve arrangements with counterbalance valves in the return is their great tendency towards unstable behavior.

The document US 5,878,647 shows a valve arrangement with two inlet and two return valves, wherein each of the valves is controllable via an individual pilot valve. Although each control valve can be controlled individually in this valve arrangement, but the device-technical effort for a total of four pilot valves is high.

In contrast, the invention has for its object to provide a device-technically simplified hydraulic valve assembly with resolved control edges for the supply and return. Furthermore, the invention has for its object to provide a proper valve block and the proper hydraulic drive.

The first object is achieved by a hydraulic valve arrangement with the

Characteristics of claim 1, the second object by a valve block with the Characteristics of claim 12 and the third object by a drive with the

Features of claim 15.

Advantageous developments of the hydraulic valve assemblies are in the

Claims 2 to 11 described. The developments of the valve block are described in the claims 13 and 14.

A hydraulic valve arrangement for the pressure medium supply of a hydraulic consumer, in particular a hydraulic cylinder has two working connections to

Connection with the consumer. In particular, the valve assembly is used in the field of mobile machines, for example in an excavator or backhoe loader, a tractor or a forklift. The valve arrangement has a first inlet orifice plate, via which a first feed flow from the first of the working connections to the consumer can be controlled. In addition, it has a separate, second return orifice, via which a first return flow from the consumer via the second working port is controllable. Thus, a first direction of movement of the consumer is defined via the first inlet flow and the first return flow. Furthermore, the valve arrangement has a second

An inlet orifice plate, via which a second feed stream from the second of the working ports to the consumer is controllable, and a separate, first return orifice, via which a second return flow from the consumer via the first working port is controllable. About the second inlet flow and the second return flow is thus a second

Direction of movement of the consumer defined. According to the invention, two separate pilot valves are provided in the valve arrangement. The pilot valves referred to as separate are characterized in particular by independent control signals. In this case, the first return flow orifice can be controlled via a first of the pilot valves to control the second return flow. To control the second feed stream is the second

Inlet meter according to the invention at the same time vorsteubar via the first pilot valve and the second pilot valve. In this case, control signals of the two pilot valves can act on the second inlet valve with the same effect or counteracting each other.

In this way, the second feed stream and the second return stream

be controlled independently of each other, about which adapted to the respective hydraulic consumer detuning the second inlet orifice plate to the first Return flow orifice, in particular the detuning of the throttle cross sections, is possible. The detuning can be applied for example by software technology. Finally, this application can essentially be automated, without having to generate the multiplicity of slide variants during startup, as is disadvantageously the case in the prior art.

In a particularly preferred development, the second return flow orifice is controllable via the second pilot valve for controlling the first return flow and, moreover, for controlling the first feed flow, the first flow orifice is simultaneously controllable via the first pilot valve and the second pilot valve. In this development, therefore, only two pilot valves are needed to the two inlet orifices depending on

To control direction of movement regardless of their associated return orifice plates. Thus, the previously discussed detuning is possible for both directions of movement of the consumer. This represents a considerable technical advantage over the prior art, since not a separate pilot valve is required for each inlet and return orifice, but the amount of pilot valves is halved.

The pilot valves are, for example, each electromagnetically actuated.

In this case, preferably a first pilot pressure can be generated from the first pilot control valve via which the first return flow orifice and the second inlet orifice are controllable. From the second pilot valve, a second control pressure is preferably generated, via which the second return flow orifice and the first inlet orifice can be controlled. Unlike the prior art thus the return flow orifices are not indirectly from the pressure of

Inlet flow orifice, but actuated directly from the associated pilot valves. The control of the return orifice plates is thus more direct and precise.

In a further development, a main control spool forming the respectively controlled inlet orifice is subjected to both control pressures.

In addition to this, a control slide having the respective controlled return flow meter can be acted upon solely with the higher or only with the lower of the two control pressures. Controlled in this context is to be understood that a throttle cross-section of the inlet orifice, or return orifice plate for measuring the inlet flow, or

Return flow is controlled.

In a development, the valve arrangement has a high-pressure connection for connection to a pressure medium source and a low-pressure connection for connection to a

Pressure medium sink. The first feed flow from the high pressure port to the first working port and the first return flow from the second working port flow to

Low pressure port. Conversely, the second feed flow from the high pressure port to the second working port and the second return flow from the first working port to the low pressure port.

In a development of the valve arrangement, the latter has a selector valve designed in particular as a directional control valve for controlling the direction of movement. In a first position of the selector valve assigned to a first direction of movement of the consumer, a measuring surface of the first main control valve having the first inlet control orifice can be acted upon by the second control pressure and a measurement surface of the second main control orifice having the second inlet orifice can be subjected to low pressure or tank pressure. In a second position of the selector valve assigned to a second movement direction of the consumer, a measuring surface of the second main control slide having the second inlet measuring diaphragm can be acted upon by the first control pressure and a measuring surface of the first main control slide having low pressure or tank pressure can be acted upon.

If, in a further development, the selector valve has a valve body against which the control pressures act counteracting one another, then the first position and the second position are set as a function of the control pressures. In this way, the direction of movement of the consumer can be determined via the control pressures.

In a development, the selector valve is formed via a directional control valve, in particular a 5/3-way valve. In an alternative development, the valve arrangement has a shuttle valve for determining a higher of the two control pressures and both the inlet orifices

forming main spool are subjected to the higher of the control pressures.

Preferably, the main spool are acted upon with the higher of the control pressures in the direction of an opening of their respective inlet orifice plate.

In a further advantageous alternative development, the two have the

Inlet flow meter forming main spool with a respective end face to each other. In this case, a piston bore is preferably introduced from the respective end faces which are assigned to one another. In the respective piston bore, a respective pressure piston can be guided. These pressure pistons may preferably define a control space with their end face in the respective piston bore. Furthermore, the pressure pistons with their mutually facing end faces can be applied to each other for common axial displacement. Furthermore, it is preferably provided that a respective pressure piston can be applied to it at a displacement in the direction of its spool. In a system of the respective pressure piston to its spool then the limited of this pressure piston control chamber is preferably closed. This solution has the advantage that in device-technically simple way both

Main spool be charged from each highest control pressure by the control chamber is zusteueuert with the lower control pressure from the pressure piston. About the control chamber with the higher pressure then the one main spool on the pressure piston and the other main spool over the control chamber with pressure medium can be acted upon. Preferably, then the one control chamber with the first pilot valve and the other control chamber is connected to the second pilot valve. In a further embodiment, a control channel formed in the respective control slide can open into the bore base of a respective piston bore. The respective control channel can then be closed at a contact of the pressure piston on the spool of the pressure piston and be open in the off state of the pressure piston. As a result, a closure of the pressure-reduced control chamber is made possible in a simple manner. Furthermore, it is conceivable that the control chambers are throttled connected to the tank. In addition, it is conceivable that a respective pressure piston has a, in particular centrally formed, axial projection for closing its associated control channel on its end face facing the control channel. Furthermore, it is conceivable that the

Main spool with the piston bores and the pressure piston in a common Control valve are formed. The Applicant reserves the right to make an independent claim to a control valve with one or more of the aforementioned aspects.

In opposite, the respective inlet orifice closing direction, the first main spool is preferably with the first control pressure and the second

Main spool applied to the second control pressure. Then remains the one of the main spool with closed inlet orifice, which is acted upon in the closing direction with the higher of the control pressures. The other one then has a controlled throttle cross-section of its inlet orifice plate.

In a further development, the valve arrangement allows a load-independent control of the consumer, by having a pressure compensator, which is fluidly connected upstream of the two inlet valves. The inlet valves via the pressure compensator are fluidic with the

High pressure connection can be connected. In the opening direction, the pressure compensator is subjected to a pressure upstream of the inlet valves, whereas in the closing direction it is acted upon by a higher one of the pressures at the working connections or a pressure dependent thereon, which is present between the working connections and the respective inlet measuring orifice. Alternatively, the pressure compensator is acted upon in the closing direction with a higher of the control pressures.

Alternatively, it is conceivable that the hydraulic valve arrangement has a respective individual pressure balance downstream of a respective metering orifice. As a result, the hydraulic valve arrangement can be used for a LUD V control. By the

Individual pressure compensators can be so much throttled between the orifices and the load of the fluid flow that a pressure to all metering orifices, preferably equal to the highest load pressure or slightly above it. This changes with a

Low supply to the pressure downstream of the orifices nothing. In front of all orifices in this case a pump pressure increases in the same way, so that the pressure difference changes in the same way at all metering orifices, while in the case of an undersupply, the pump pressure decreases, and the current distribution between the metering orifices is maintained.

In a further development, in addition, a pressure-selection valve is provided, via which the higher of the pressures at the working ports, or the pressure dependent thereon, against the higher the control pressures can be weighed, and the pressure compensator is acted upon with the thus weighed highest of the pressures in the closing direction, cavitation in the respective inlet can be effectively prevented. In this way, at a falling inlet pressure, especially in active or so-called pulling loads no longer the inlet pressure but the highest of the control pressures is reported to the pressure compensator. In this way, an active refilling can be realized, which prevents cavitation as already mentioned.

For the function of active refilling, the valve arrangement preferably has a suction valve, via which the low-pressure connection can be fluidically connected to the respective inlet. Particularly preferred in this Nachsaugventil in addition to a

Pressure relief valve function integrated, so that at the same time the inlet is protected against overload. Preferably, both feeds, that is, the respective

Pressure medium channels between the inlet orifices and the working ports, such a Nachsaug- and load protection function.

In a preferred development, the valve arrangement has a

Regenerationswegeventil via which the first return flow is at least partially fed to the first feed stream. This development proves to be advantageous if, for example, for reasons of acting gravity pressure medium is displaced from the hydraulic consumer. If the consumer is a hydraulic cylinder and moves, for example, a handle of an excavator, so can be displaced from the hydraulic cylinder in the return when lowering the stem pressure and fed directly to the inlet via the regeneration valve. In this way, a power consumption of the pressure medium source, in particular a hydraulic pump, can be reduced by the corresponding amount of the regeneration volume flow via the regeneration directional control valve.

In addition, the second return flow can be at least partially supplied to the second feed stream via said or an additional regeneration directional control valve.

Preferably, the regeneration directional control valve has a regeneration position in which the return line can be connected to a pressure medium line, via which the pressure compensator can be connected to the respective inlet orifice. Preferably, the Renerationswegeventil a second position, via which the respective return is connectable to the low pressure port.

In the case of a hydraulic consumer configured as a differential cylinder, the regeneration directional control valve additionally has a regeneration position in said regeneration position

Pressure medium connection of the return to the low-pressure connection, if the return is connected to that of the pressure chambers of the hydraulic cylinder having the larger Kolbenfiäche. As in the associated direction of movement of the consumer (displacement of pressure medium from the piston chamber in the return) of the inlet into the annulus of the

Hydraulic cylinder is smaller than the return from the piston chamber, there is a positive flow surplus in the return, via the Renerationswegeventil to

Low pressure connection must be dissipated. For the alternative case, in which pressure medium is displaced from the annular space of the hydraulic cylinder in the return at active load, it is sufficient if the regeneration directional valve said two positions - pressure fluid connection of the return to the inlet or pressure medium connection of the return to the

Low pressure connection - has. There must be no simultaneous splitting of the return flow to the inlet and the low pressure connection.

In a preferred development, the valve arrangement has a control unit in which a characteristic map of the first control pressure is stored as a function of the second control pressure, and / or vice versa. Depending on the development, the detuning of the second inlet orifice plate for the first return flow orifice plate for the second direction of movement and / or the detuning of the first orifice orifice plate for the second return flow orifice for the first direction of movement is then stored in this characteristic map.

In electromagnetic actuation of the pilot valves is in the control unit preferably a map of the first control pressure in response to a first

Actuating current or a first actuating voltage and / or a map of the second control pressure in response to a second operating current or a second

Actuating voltage stored. Preferably, at least one of the pilot valves and / or one of

Feed orifices and / or the Rücklaufmessblenden and / or the regeneration valve designed in slide construction.

In terms of apparatus technology, the valve arrangement proves, in a further development, in which the first inlet orifice and the first return orifice are jointly formed by a first main valve body. Alternatively or additionally, the second

Inlet orifice and the second return orifice together by a second

Main valve body be formed.

Preferably, the valve assembly, which is designed according to at least one of the preceding aspects, combined in a hydraulic valve block. This results in a particularly compact and robust valve arrangement. Preferably, the

Valve block a plurality of valve assemblies, wherein each valve assembly is then preferably associated with a valve disc. The valve discs are preferably in stacks

joined together and bordered by an end plate and a connection plate. Preferably, the connection plate and the valve disks pass through a high-pressure channel

Low pressure channel and optionally a load signaling channel. In this case, the high-pressure channel and the low-pressure channel respectively connect the said high-pressure connections and

Low pressure connections of the respective valve arrangement, or valve disc. In the load reporting channel of each valve assembly of the load pressure of the supplied by her hydraulic consumer is preferably reported via a shuttle valve or a Wechselventilkaskade. Thus, in the load reporting channel, the highest of the load pressures of the consumer.

Manufacturing and device technology is simple, the valve block proves in a development in the valve spool of Zulaufmessblenden and / or the

Return flow orifices are arranged together in a valve bore.

A similar production engineering and device technical advantage uses a development in which a valve spool of the pressure compensator and a valve spool of the

Regenerationswegeventils are arranged in a valve bore together. A hydraulic drive has a, in particular adjustable hydraulic pump for

Pressure medium supply at least one hydraulic consumer, in particular

Hydraulic cylinder. In addition, the hydraulic drive has a valve arrangement which is formed at least according to one of the aspects of the preceding description. The valve assembly is preferably arranged in a valve block, as it emerges from the description.

Nine embodiments of a valve arrangement according to the invention, a

Embodiment of a valve block according to the invention and an embodiment of a drive according to the invention are shown in the drawings. With reference to the figures of these drawings, the invention will now be explained in more detail.

Show it

Figure 1 is a hydraulic circuit diagram of a hydraulic drive according to the invention, Figures 2 to 10 and 16 embodiments of inventive valve assemblies, Figure 11 is a longitudinal section of a hydraulic valve block according to the invention with a valve assembly according to the invention according to Figure 7 and

Figures 12 to 15 in a circuit diagram or a longitudinal section of embodiments of

Main spools.

Figure 1 shows a hydraulic drive 1 according to the invention with two hydraulic consumers 2, which are each designed as a differential cylinder. Furthermore, the hydraulic drive 1 has a hydraulic pump 4 with adjustable displacement. This works in the open hydraulic circuit and is connected with its low-pressure connection with a pressure medium sink, which is designed as a tank T. From there it promotes pressure medium in a working line 6, via which the hydraulic consumers 2 can be supplied with pressure medium.

Each of the hydraulic consumers 2 is assigned a valve disk 8 of a valve block 10. The valve block 10 has a high pressure port P, a low pressure port T and a load reporting port LS. On one side beyond the ports P, T, LS, the last valve disc 8 is closed with an end plate 12 of the valve block 10. The hydraulic pump 4 is assigned a pressure regulator 14 to which the highest of the load pressures of the hydraulic consumers 2 is reported via the load reporting connection LS. The designed as a 3/2-way valve pressure regulator 14 is in the direction of a Durchfiussstellung in which the displacement volume of the hydraulic pump 4 is increased over it, with the pressure in the

Working line 6, and in a flow-through position, in which the displacement volume is reduced via it, loaded with the pressure at the load-signaling connection LS and a pressure equivalent of a control spring. About the pressure regulator 14 is a working space of an adjusting unit 16 for adjusting the displacement volume of the hydraulic pump 4 on the one hand with the working line 6 and on the other hand with the tank T fluidly connectable. Via the pressure regulator 14, a pressure in the working line 6 is thus provided, which is higher than the highest load pressure by the pressure equivalent of the spring of the pressure regulator 14.

The valve discs 8 each have one of the valve arrangements according to the invention, as shown for example in Figures 2 to 10. An exemplary embodiment of such a valve arrangement, as corresponds to that according to FIG. 7, is shown in a longitudinal section of the valve disk 8 according to FIG. For a basic understanding of

Valve assemblies according to the invention, however, first Figures 2 to 10

described.

According to FIG. 2, a valve arrangement 18 has a high-pressure connection P a

Low-pressure connection T and a load-signaling connection LS. As already shown in FIG. 1, the valve arrangement 18 is arranged in a valve disk 8, which is symbolized in FIG. 2 by means of a dot-dash line. The high pressure port P is connected via the working line 6, which passes through the valve disc 8 to the next valve disc 8 according to Figure 1, with the high pressure port P of the following valve disc 8. Accordingly, the

Low pressure port T via a low pressure line 20 to the low pressure port T of the subsequent valve disc 8 is connected. From the high pressure line 6 branches one

High pressure line 22 from which is connected to a 3/2-pressure compensator 24. This is connected to a pressure line 26, branch off from the two pressure branches 28 and 30. A first pressure branch 28 is connected to a first inlet valve 32 and a second pressure branch 30 to a second inlet valve 34. The first inlet valve 32 is via a first

Working line 36 with a first working port A of the valve assembly 18, and the valve disc 8, fluidly connectable. The second inlet valve 34 is in the same way via a second working line 38 with a second working port B of the valve assembly 18, and the valve disc 8, connectable. According to FIG. 1, the first working connection A is connected to a piston chamber 40 and the second working connection B is connected to an annular space 42 of the hydraulic consumer 2.

Furthermore, the valve arrangement 18 has a first pilot valve 44 and a second pilot valve 46. Both are designed as 3/2 proportional directional control valves and can be actuated in each case via an electromagnet 48 or 50.

Furthermore, the valve arrangement 18 has a first return valve 52 and a second return valve 54. Both return valves 52, 54 are pilot-controlled in the embodiment shown in each case via a pilot stage 56, but may deviate from it also be designed directly controlled.

Both pilot valves 44, 46 are supplied with pressure medium from the high-pressure line 22 by branching off from it control pressure lines 58.

The two load-signaling connections LS shown in FIG. 2 of the adjacent valve disks 8 according to FIG. 1 are connected to a shuttle valve 60. The output of the shuttle valve 60 is depending on the number of valve disks 8 either a subsequent

It is connected to the load-signaling connection LS of the valve block 10 according to FIG. 1 and consequently reports the highest of all the load pressures of the valve arrangements 18 to the pressure regulator 14. The second input of the change-over valve 60 is a load-reporting line 62. which is connected to an output of a further shuttle valve 64, which is connected via one of its inputs to the first inlet valve 32 and with the other of its inputs to the second inlet valve 34. The inlet valves 32, 34 each have an inlet position b, in which the respective pressure branch 28 or 30 with the corresponding working line 36, 38 via a

Inlet metering or 66 is connected. Downstream of the respective inlet orifice 66, 68, the load pressure prevailing there is picked up and reported to the load pressure connection LS of the respective inlet valve 32, 34. Both load pressures are weighed at the shuttle valve 64 and the higher is reported to the shuttle valve 60. There it will be with the correspondingly determined load pressure of the adjacent valve disc 8 compared and the higher of the two is forwarded to the load reporting port LS of the valve disc 8 shown.

The working line 36, or 38 is on the return orifice 88, or 90, a

Return position a of the inlet valve 32, or 34, via low-pressure lines 70 with the

Low pressure line 20 and connected via the low pressure port T to the tank T shown in FIG. Both inlet valves 32, 34 have a displacement sensor 72nd

Furthermore, the valve assembly 18 has a logic valve 74 through which two

Movement directions of the hydraulic consumer 2 are definable. The logic valve 74 is designed as a 5/3-way switching valve. A first control input 76 is connected to a first control line 78, which is connected to a pressure medium output of the first pilot valve 44. A second control input 80 of the logic valve 74 is connected to a second

Control line 82 is connected, which is connected to a pressure medium output of the second pilot valve 46. A first control output 84 of the logic valve 74 is connected to a measuring surface of the first inlet valve 32, a second control output 86 is connected to a measuring surface of the second inlet valve 34. The one connected to the first control output 84

Measuring surface of the first inlet valve 32 counteracts the pressure in the first control line 78, the counter to the second control output 86 connected measuring surface of the second inlet valve 34 counteracts the control pressure in the second control line 82. A tank connection of the

Logic valve 74 is connected via the low pressure line 70 to the low pressure line 20.

In the first control line 78, via the first pilot valve 44, proportional to an electromagnetic drive signal of the first electromagnet 48, a first

Control pressure to be generated. Accordingly, a second control pressure can be generated in proportion to an electromagnetic drive signal of the second electromagnet 50 via the second pilot valve 46 in the second control line 82.

If the piston of the hydraulic consumer 2 according to FIG. 1 is now to be moved in such a way that a second direction of movement in FIG. 1 results from right to left, pressure medium must be supplied to the annular space 42 via the second working port B and from the piston chamber 40 via the first working port A pressure medium to be discharged. For this purpose, the first pilot valve 44 is energized with a first control current stronger than the second Pilot valve 46 with a second control current. Accordingly, in the first control pressure line 78, a first control pressure is established, which is higher than a second control pressure, which is set in the second control pressure line 82. Due to this, the logic valve 74 is moved from its neutral position a to its second position b. This will be the first one

Control output 84 to the tank port T and the second control output 86 to the first control input 76 is connected. Accordingly, the first inlet valve 32 remains in its return position a and the second inlet valve 34 is actuated in its inlet position b.

The pilot stage 56 of the first return valve 52 adjusted due to the self-adjusting force balance of the pressure force of the first control pressure in the first

Control pressure line 78 and a counteracting spring force their position. Accordingly, a main stage of the first return valve 52 follows the position of the pilot stage 56 due to a setting of the relative position of the two slides of the valves 56, 52

Aperture cross-section, which relieves a spring chamber of the main stage 52 and thus ensures a pressure drop. This control corresponds to the principle of the sequential control, which is known from the prior art, for example, the EHR 23 valve of the applicant.

Due to the present at the pilot stage 56 of the second return valve 54, lower second control pressure, the main stage of the second return valve 54 receives its inlet position, in the pressure medium due to the non-return function of the main stage of the second

Supply valve 34 can flow toward the second working port B, but not vice versa.

With strong energization of the first control valve 44 and less energization of the second control valve 46 is thus the second feed stream via the second working port B in the annular space 42 and the second return flow from the piston chamber 40 via the first

Working port A allows.

According to the invention, the feed flow is controlled via the two control pressures of the first control pressure line 78 and the second control pressure line 82 acting on the second inlet valve 34. According to the balance of power from the second inlet valve 34th

acting spring and a pressure difference of the second inlet valve 34 attacking, counteracting control pressures is then a corresponding throttle cross-section of the inlet orifice 68 a. The first return valve 52 is due to compared with the second control pressure higher first control pressure switched to its return position and releases the first working line 36 for the second return flow. Since the higher first control pressure acts on the first inlet valve 32 in the direction of the return position a, and only the tank pressure is effective in the direction of the inlet position b, the first inlet valve 32 is switched into the return position a with the assistance of the spring force. Accordingly, the second return flow in the first working line 36, the first return orifice plate 88, the return position a of the first inlet valve 32 to the low pressure line 70 in

Essentially unthrottled. The second feed stream via the second working port B, however, as described, depending on the second inlet valve 34 acting

Control pressures and the pressure equivalent of the spring throttled. As a consequence, the two throttle cross sections, the inlet flow and the return flow are set independently of each other.

To the hydraulic consumer 2 now in the opposite, first

Moving direction, the control of the pilot valves 44, 46 must be reversed. Then, the second pilot valve 46 outputs the higher second control pressure, and the first pilot valve 44 outputs the lower first control pressure. Accordingly, the first inlet valve 34 moves into its inlet position b and the second inlet valve 34 moves in its return position a. As already described in the second direction of movement, due to the control pressure conditions, the main stage of the first return valve 52 remains in its spring-loaded position, and allows the inlet to the first working port A, whereas the main stage of the second return valve 54 is pushed by its pilot stage 56 in the return position. Accordingly, a first feed flow via the first inlet valve 32 and the first working port A and a first return flow from the second working port B via the main stage of the second return valve 54 and the

Set return position a of the second inlet valve 34. The main spool 92, 94 of Figure 2 are equipped with the displacement sensors 72, which is particularly necessary when used in the tractor. About the control of the slide position is a very accurate adjustment of the aperture cross section of the respective inlet orifice plate. Thus, in

Connection with the regulation of the pressure difference across the respective inlet orifice a very accurate inlet flow rate can be adjusted. FIG. 3 shows a second exemplary embodiment of a hydraulic valve arrangement 18, which is modified in the region of the inlet valves 32, 34 and the return valves 52, 54 compared to the first exemplary embodiment. In summary, it can be said that the inlet orifice 66 of the first inlet valve 32 and a return orifice 88 of the first return valve 52 are formed together on the same main spool 92.

Accordingly, the metering orifice 68 of the second inlet valve 34 and a

Return flow meter 90 of the second return valve 54 at a common

Main spool 94 is formed. Accordingly, the main spool 92 and the main spool 94 have three positions a, b, c. In a middle position a while the working line 36, or 38 is separated from both the high pressure line 22 and from the low pressure or tank line 70. The LS connection of the valves 32, 52, and 34, 54 is relieved in the middle position a via the low-pressure line 70 to the tank. Due to the

Formation of the metering orifices 66, 88, and 68, 90 on the respective main spool 92, and 94, the blocking function of the respective return valve 52 and 54 are executed device switching simply switching. The main spool 92, 94 of Figure 3 are equipped with the displacement sensors 72, which is particularly necessary for a tractor application. As a result, by regulating the slide positions, an accuracy of the setting of the necessary area ratio of supply and return metering orifice 66, 90 of the first movement direction, or 68, 88 of the second movement direction, can be increased. Thus, a tolerance of the clamping pressure of the main spool 92, 94 decreases. In order to set a free position of the hydraulic consumer 2, the two pilot valves 44, 46 must be controlled with the same current signal value. This also applies to the previous one

described embodiment of Figure 2.

In the first two presented hydraulic control arrangements 18 according to Figures 2 and 3, the control pressure selection for setting the corresponding inlet orifice 66 and 68 via the logic valve 74. This must have five connections and three positions, resulting in a device complex and therefore expensive

Design leads. The following exemplary embodiments therefore propose a changeover valve 96 for control pressure selection and for simplifying the apparatus

Control pressure forwarding. In the following explanation of FIG. 4, only the features which differ from the preceding exemplary embodiments are explained in more detail. Also in this embodiment, the higher of the two control pressures, given by the two pilot valves 44, 46, again predetermines the positioning of the return orifice orifice 88 or 90. In the event that the first control pressure in the first control pressure line 78 is greater than the second control pressure in the second control pressure line 82, so connects in this

Embodiment newly added shuttle valve 96 in the direction of

Inlet positions b effective measuring surfaces of both inlet valves 32, 34 with the first

Control pressure line 78. acts on the first inlet valve 32 in the opposite direction, in the direction of the return position a, the higher first control pressure together with the

Pressure equivalent of a spring, so that the first inlet valve 32 remains in the return position a. Due to the second control pressure in the second control pressure line 82 being lower, the higher first signal reported by the shuttle valve 96 to the second supply valve 34 may be higher

Control pressure to adjust the valve body of the second inlet valve 34 in the inlet position b. The control of the return valves 52, 54 is analogous to the first embodiment shown in FIG 3 via the pilot stages 56. Thus, in terms of device technology simpler way, the second feed flow via the second inlet valve 34 toward the second working port B and the second return flow through the first working port A, the first return valve 52 and the first inlet valve 32 a. The advantage of the described valve arrangement 18 according to FIG. 4 lies above all in the less complex and therefore more robust construction, since the logic valve 74 according to FIGS. 2 and 3 is eliminated and thus no additional control notches and channels are to be provided for the control pressures to be selected. In order to adjust the clearance position of the hydraulic consumer 2, the two pilot valves 44, 46 must be controlled as already mentioned with the same current signal value.

The exemplary embodiment of a valve arrangement 18 according to FIG. 5 is based essentially on that of FIG. 4, and likewise has the shuttle valve 96 instead of the logic valve 74. Thus, the advantages of the simplified control pressure selection by the simple shuttle valve 96 and the common design of the first metering orifice 66 and the first return orifice 88 at the first main spool 92, and the second metering orifice 68 and the second return orifice 90 at the second main spool 94 (see Figure 3). combined. In order to adjust the clearance position of the hydraulic consumer 2, here again the two pilot valves 44, 46 must have the same current signal value be controlled. An advantage of this embodiment over the previous

Valve arrangements, represents the possibility by adapting to the

Main control valves 92 and 94 effective measuring surfaces and their differences in a later design to ensure that the throttle cross-section of each driven

Inflow orifice 66, and 68 always (significantly) fails greater than the throttle cross-section of each driven return orifice 90, or 88. On the main spool 92, and 94 of the controlled inlet orifice 66, or 68 attacking lower the

Control pressures can then according to the invention, the throttle cross-section of the inlet orifice plate 66, 68 are reduced as needed again.

Figure 6 shows a valve assembly 18, in particular for an excavator or

Backhoe loader is suitable. It has an active refill function. As already described, for a low-tolerance and manufacturing influences insensitive adjustment of the clamping pressure of the main spool 92, 94 of the inlet valves 32, 34 theoretically sensing the

Positions of these valve body necessary. Thus, the throttle cross-section of the respectively controlled inlet orifice 66, or 68 and return orifice orifice 90, or 88 can be controlled and adjusted very accurately by controlling the position. However, since this additional displacement sensors 72 are required and an evaluation and control electronics is necessary, the cost of such a valve assembly for use in excavators and backhoe loaders increase unnecessarily, since there the exact adjustment of the feed stream is not required in contrast to the application in the tractor. This regulation usually takes over the operator of the excavator or backhoe loader. For this reason, a valve assembly 18 with active refill function for use in excavators and backhoe loaders without displacement sensors 72 is proposed according to FIG. Since for excavators or backhoe loaders a tightness function usually plays no or only a minor role, can also be dispensed with a use of shut-off valves in the return. However, they are

Pressure limiting valves 98 are provided, via which the two working lines 36, 38 are each connected to the tank line 20. Advantageously, the pressure limiting valves 98 also have an integrated suction valve 100, via which the respective low pressure leading working line 36, or 38 can draw in pressure medium from the low pressure line 20 in order to prevent cavitation. In contrast to the previous exemplary embodiment, the valve arrangement 18 according to FIG. 6 has an additional shuttle valve 102, which compares the higher of the two load pressures with the higher of the two control pressures. The highest of the kind weighed pressures is then reported to the individual pressure compensator 24 and the shuttle valve 60. Thus, an active refilling is realized by the fact that with decreasing inlet pressure, for example, active or pulling loads, not the inlet pressure, but the maximum control pressure to the LS port of the valve assembly 18 is reported. The cavitation in the inlet channel can be prevented, since in this way the pressure between the individual pressure compensator 24 and the respective inlet orifice 32 or 34 is increased.

In particular, for the applications in excavators or backhoe loaders, a regeneration function should be provided to increase the energy efficiency for some consumers, such as the hydraulic cylinder that articulates the handle of the excavator / backhoe loader. The

Valve assembly 18 of FIG. 7 providing it is substantially based on that of FIG. 6, but includes an additional regeneration directional valve 104. This is designed as a 3/2-way valve and has two pressure medium outputs, one of which with the

Low pressure line 20 and the other is connected to the pressure line 26. One

Pressure medium input of the regeneration path valve 104 is connected to a return outlet 106 of the integrated second supply and return valve 34. The return outlet 106 is either shut off depending on the switching position of the valve 34 or connected via the second return metering orifice 90 with the then acting as a return second working line 38. At the valve body of the regeneration path valve 104, a spring engages in the direction of a first flow position, in which the return outlet 106 is connected to the low-pressure line 20 at. Equally effective, the valve body is loaded with the pending in the pressure line 26 pressure downstream of the individual pressure compensator 24. Toward the regeneration position b of the

Regenerationswegeventils 104, the pressure in the second working line 38, or the return pressure at the second working port B or a dependent pressure acts. Now, if the pressure in the pressure line 26 is higher than the pressure in the second working line 38, so is a passive, to be moved load on the consumer 2, and the valve body of

Regenerationswegeventils 140 remains in the basic position a. The return flow rate through the second working line 38 can thus flow completely into the low-pressure line 20. However, if there is an active or pulling load on the hydraulic consumer 2, the pressure in the second working line 38 is greater than the pressure in the pressure line 26, so that the regeneration path valve 104 assumes its regeneration position b. As a result, the pressure medium volume flow that is completely displaced from the annular space 42 flows via the

Regenerationswegeventil 104 completely back into the pressure line 26, whereby the hydraulic pump is relieved to provide the inflow volume flow. The valve arrangement shown in Figure 7 is particularly suitable for excavators or backhoe loaders, as there often the steering arm or the boom articulating hydraulic consumer 2 is arranged so that pulling loads act mainly in the direction of extending piston. Therefore, the pressure in the annular space 42, or the dependent pressure in the second working line 38 is used for the circuit of the regeneration path valve 104 in the regeneration position b.

Compared to the last-mentioned embodiment, the embodiment according to FIG. 8 has a modified regeneration-way valve 104. The reason for this is that the hydraulic consumer 2 is in the reverse installed position, whereby the pulling or active load enters the piston and not, as in the embodiment described above, extends. As a result, the piston chamber 40 is reduced in pulling load, whereby the return flow is greater than the required inlet flow. As a result, part of the

Return flow volume towards the tank T are discharged via the low-pressure line 20, even if the regeneration position b of the regeneration valve 104 is activated. Only a portion of the return flow from the piston chamber 40 thus returns to the inlet, d. h., In the pressure line 26. For this purpose, the regeneration position b of the regeneration control valve 104 in this embodiment, a pressure medium connection of the return outlet 106 of the integrated supply and return valve 34 with both the pressure line 26, and with the low-pressure line 20.

FIG. 9 shows an exemplary embodiment of a valve arrangement 18 which is based on that of FIG. Deviating from this, however, it has no LS interconnection, which is why the individual pressure compensator 24 and the shuttle valves 64, 102 and 60 according to FIG. 7 are missing. Correspondingly simpler and the main spool 92, 94 are formed at the respective center position, the working line 36, or 38 is shut off only against the high pressure line 6 and the low pressure line 20. Since the valve arrangement 18 according to FIG. 10 also has a regeneration function, it also has the regeneration directional valve 104, but here via the regeneration position b of the valve 104, the regeneration output 106 of the integrated supply and return valve 34 directly to the first working line 36 downstream of the metering orifice 66 connected is. With the exemplary embodiment of the valve arrangement 18 according to FIG. 9, for example, it is also possible to operate the valve in a future-proof manner in a fully electronified environment. When using suitable sensors (eg pressure sensors in the load ports, position sensors on the hydraulic cylinders) and the possibility of supply and

Separate return flow orifices can be adjusted separately, with the help of a suitable operating and control strategy, a performance can be adjusted so that the clamping of the pressures can be reduced to a minimum. Possibly. can be achieved with such controls also a higher degree of regeneration.

All previously presented valve assemblies 18 have the common feature that the respective return flow is set purely controlled. The advantage of this

Valve arrangements 18 are limited to the already discussed, easier commissioning by the possibility of automated application and the software-related, independent setting of the throttle cross-sections of the metering orifice 66, and 68 and the associated return orifice plate 88, or 90. In addition, the variance decreases

Valve shifters of the inlet and return valves in the manufacturing process, as the variance is shifted into the software. However, the setting of the respective return flow rate at purely controlled return orifice plate 88, or 90 depends on the load pressure in each return channel 36, or 38 dependent. This makes a relatively high clamping of the corresponding inlet pressure necessary, which reduces energy efficiency of the hydraulic valve assemblies 18, and the hydraulic drive 1.

The valve arrangement 18 according to FIG. 10 therefore proposes a solution in which the pressure upstream of the respective return flow orifice 88, or 90, via a pressure regulating valve arrangement 108 for the first working line 36 acting as return and via a pressure regulating valve arrangement 110 for the second working line 38 functioning as return is controllable. This results in a dependent of the respective control pressure pressure difference across the respective return orifice plate 88, or 90. As a consequence, in addition to the inlet flow, now also the respective return flow is substantially only from

Throttle cross section of the respective return orifice plate 88, or 90, and no longer dependent on the load applied to the return pressure. For this purpose, the respective

Pressure control valve assembly 108, and 110 a poppet 112, and 114, which is actuated by the first control pressure, and the second control pressure. The respective topping piston 112, and 114 actuates its associated pilot stage 56, wherein the pressure in the first working line 36 upstream of the first Rücklaufmessblende 88 with the first control pressure on the poppet piston 12 and the pressure in the second working line 38 upstream of the second Rücklaufmessblende 90 on poppet 114th is compared with the second control pressure. As a result, a regulation of the respective pressure before the return flow orifices 88, or 90, which depends on the respective control pressure of the pilot valve 44, or 46.

Are the dimensions of the hydraulic consumer 2, in particular the

Area ratios on the piston, known, the return orifice plate 88, or 90 can be tuned in each operating range so that one with respect to the associated

Inlet volume flow cylinder area ratio adjusted, slightly larger

Return flow follows. In this way, the condition that the supply pressure is not clamped, is met. In turn, the higher of the two pressures from the maximum control pressure and the inlet pressure can be selected via the shuttle valve 102 and reported to the spring-loaded end face of the individual pressure compensator 24. If the inlet pressure drops below the value of the maximum control pressure with active (pulling) loads, the pressure in channel 26 is no longer regulated instead of the pressure difference across the active or controlled inlet orifice plate. The pressure gradient across the active or controlled inlet orifice increases and thus cavitation can be avoided by active refilling.

FIG. 11 shows by way of example a possible embodiment of a valve block, more specifically a valve disk 8, which has a valve arrangement 18 according to the exemplary embodiment from FIG. It is conspicuous that a valve spool of the regeneration path valve 104 together with a valve spool of the individual pressure compensator 24 together in a

Valve bore 116 of the valve disc 8 is received. Furthermore, the two

Main spool 92, 94 with the metering orifices 66, 68 and

Return flow orifices 88, 90 are also arranged in a common valve bore 118. Both have a favorable effect on the production and the cost of the valve disc 8. The weighed in the shuttle valve 96, higher of the two control pressures is directed to a trained between the main spools 92, 94 pressure chamber 120. The latter is bounded on both sides by a respective piston 122, or 124, on which the measuring surface acting in the direction of the inlet position c is arranged. In the following figures 12 to 14, two advantageous extensions of

Control of the two inlet valves 32 and 34 presented with their associated inlet orifices 66 and 68. Common to both is the fact that it is possible to dispense with the shuttle valve 96 for selecting the highest control pressure (compare FIGS. 4 to 11). Instead, two check valves 122, 124 are used. The advantage lies in the favorable placement of the check valves 122, 124 within the main spool 32 and 34. The structural design is shown in FIG 14. This eliminates the processing of another axis

(Shuttle valve 96), which leads to lower costs in the manufacturing process. Furthermore, the assembly for adjusting the differential surface of the main control valves 92, 94 can be much simpler and thus probably cost less expensive. This concept requires that the pressure chamber 120 must be relieved via a, in particular small, aperture to the tank T. Leaks from the control pressure are thus the result. Furthermore, it is no longer possible to fall back on the maximum control pressure for the active refill, unless a separate shuttle valve is used again. However, the advantages mentioned in the simpler structure of this unit of the main spool 92, 94 continue to exist.

Figure 12 shows a part of the just described control of the

Main spool 92, 94, taking care to elaborate the details of each

Switch positions has been omitted. In turn, the higher of the two control pressures from the control lines 78 and 82 into the pressure chamber 120 can be reported via the check valves 122 and 124. So that the check valves 122, 124 can open and the pressure in the pressure chamber 120 can degrade, an orifice 126 is provided from the pressure chamber 120 to the tank T. The adjustment of the main spool 92, 94 takes place as in the

previously described figures 4 to 10.

A particularly favorable elaboration of a possible control, the figures 13 and 14. For example, if the first control pressure of the first control line 78, the higher of the two control pressures, builds in a pressure chamber 128 (bore in the first

Hauptsteuerschieber 92) in about the required control pressure. However, leakage through an orifice 136 reduces the amount of control pressure to be adjusted to some extent. This pushes a piston 132 on a piston 134, whereby the latter is driven in its stop on Hauptsteuerschieber 94. In this case, pressure medium through an aperture 138, which is shown in Figure 4 as a slot in the piston 134, escape to the tank T. The Check valve 124 remains closed. The second main spool 94 is now mechanically displaced by the pistons 132 and 134. The compressive force is due to the

Control pressure generated in the pressure chamber 128. For example, the inlet orifice plate on the main spool 94 to be able to adjust independently, the control pressure in the control line 82 must be adjusted, whereby he always this one by the ratio of differential area and

End face on a spring assembly must be smaller than the control pressure in the control line 78. The orifices 138 and 136 in turn serve to relieve the pressure chambers 130 and 128. If the hydraulic consumer 2 are moved in the other direction of movement, the second control pressure in the control line 82 must be greater than the first control pressure in the

Control line 78. On the other hand, for the release circuit both control pressures must assume the same values.

A further advantageous embodiment is shown in FIG. 15, in contrast to the embodiment in FIGS. 13 and 14 no check valves are provided, but the control lines 78 and 82 (control channels) can be controlled and controlled via the pistons 132 and 134. The pistons 132 and 134 (pressure piston or differential piston) are guided in a respective piston bore 140, 142 of a respective main spool 92, 94. In the bore bottom 144, 146 of a respective piston bore 140, 142 opens approximately centrally the respective control line 78, 82. Furthermore, a respective piston 132, 134 on its bore bottom 144, 146 facing end side 152, 154 has a sealing geometry in the form of an axial projection 156, 158. The axial projections 156 and 158 are each approximately conical or frusto-conical in shape, wherein they taper in the direction of their associated control line 78 and 82, respectively. With the axial projection 156 of the piston 132 is then the

Control line 78 and with the axial projection 158 of the piston 134, the control line 82 closed. The control lines 78 and 82 can completely pass through the respective main valve spool 92, 94. The pistons 132 and 134 are with their facing end faces 160 and 162 abutted against each other, whereby both pistons 132, 134 are axially displaceable together. Via the first control line 78, the end face 152 of the piston 132 and via the control line 82, the end face 154 of the piston 134 can be acted upon in each case with pressure medium. The larger of the thereby acting on the end surfaces 152 and 154

Control pressure shifts the pistons 132, 134 in the direction opposite to the lower control pressure. If, for example, the control pressure acting on the end face 152 is higher, the pistons 132 and 134 are moved in the direction of the second control line 82 until they reach the axial projection 158 of the piston 134 is closed. The control pressure acting on the end face 152 then acts on the second main spool 94 via the pistons 132 and 134, since the piston 134 abuts against the latter. Thus, the main spool 94 is moved axially along with the pistons 132 and 134. About a non-illustrated in Figure 15 end face of the second main spool 94, this can with control pressure of the second

Control line 82 are acted upon in the direction of the first main spool 92. This allows the second main spool 94 regardless of the position of the first

Main spool 92 are adjusted, since the control pressure of the second control line 82 is independent of the control pressure of the first control line 78. With the described

Furthermore, the first main spool 92 is pressurized with the control pressure of the first control line 78 via the bore bottom 144 with control pressure in the direction away from the second main spool 94. If the control pressure of the second control line 82 is greater than the control pressure of the first control line 78, the pistons 132, 134 are moved to the first main spool 92, whereby the first control line 78 is closed and the first main spool 92 via the pistons 132 and 134 with control pressure the second control line 82 is acted upon. Furthermore, the second main spool 94 is acted upon by the bore bottom 146 with the control pressure of the second control line 82 in a direction away from the first main spool 92. For clearance, the control pressures of the control lines 78, 82 are the same. The main spools 92 and 94 can be adjusted independently of the position of the other main spool. Furthermore, facing each other annular end surfaces 164, 166 of the main spool 92, 94 and the end faces 160, 162 of the piston 132, 134 relieved to the tank.

In the embodiment in FIG. 15, no check valves are required in the main spool 92, 94 as compared to the embodiment in FIG. Furthermore, no shuttle valve is necessary in comparison to previous embodiments for selecting the higher control pressure. Moreover, in the embodiment according to FIG. 15, there is no leakage due to the control pressure in comparison to the embodiment according to FIG. 14 with the check valves, since no relief grooves 136 and 138 are necessary. The

Embodiment according to Figure 15 leads to a simple, robust and cost-effective design of a main valve axis. Referring to Figure 16, the use of the first and second main spools 92 and 94 in a LUDV arrangement is shown. The control of the main spool 92 and 94 takes place approximately in accordance with the embodiment in Figure 12. The main spools 92 and 94 are the pilot valves 44, 46 assigned. Furthermore, the first main spool 92 is followed by a first individual pressure compensator 168 and the second main spool 94 is followed by a second individual pressure compensator 170.

The main spool 92 controls a pressure medium connection between a

Tank connection T and a working port AI, which is connected via a working line 172 to the working port A. Furthermore, the main spool 92 controls a pressure medium connection between the pressure port P and a working port A2, which is connected via the individual pressure compensator 168 to the working port A. In the central positions a of the main spool 92, the ports AI, A2, T and P are separated from each other. In the positions b, the working connection A is connected to the tank connection T via the working line 172 and the pressure connection P is disconnected from the working connection A2. In the positions c of the main slide 92, however, the pressure port P is connected to the working port A2 and the tank port T is disconnected from the working port AI. Corresponding connections and positions also has the second main spool 94. Upstream of the pressure port P, both the first main spool 92 and the second main spool 94 each have a check valve 174, 176 associated therewith, each opening in the direction of flow to the respective main spool 92 and 94.

Disclosed is a hydraulic valve arrangement for supplying pressure medium to a hydraulic consumer with two directions of movement, the speed of each of which via an inlet orifice plate and one of these, separately controllable

Return flow meter are controllable. According to the invention, two separate pilot valves are provided, one of the return flow orifices being controllable via one of the pilot valves and the inlet orifice plate associated with this return orifice being simultaneously controllable via both pilot valves.

Also disclosed are a hydraulic valve block with at least one such valve arrangement and a hydraulic drive with a hydraulic pump, at least one hydraulic consumers and at least one such valve assembly or such a valve block.

List of reference numbers:

1 hydraulic drive

2 Hydraulic consumers

4 hydraulic pump

6 work management

8 valve disc

10 valve block

12 end plate

14 pressure regulator

16 adjustment unit

18 Hydraulic valve arrangement

20 low-pressure line

22 high pressure line

26 pressure line

28, 30 pressure branch

32 first inlet valve

34 second inlet valve

36 first working line

38 second working line

40 piston space

42 annulus

44 first pilot valve

46 second pilot valve

48 first electromagnet

50 second electromagnet

52 first return valve

54 second return valve

56 pilot level

58 control pressure line

60 shuttle valve

62 Load reporting line

64 shuttle valve first inlet orifice second inlet orifice

Low-pressure line

displacement sensors

logic valve

first control input first control line second control input second control line first control output second control output first return orifice second return flow orifice first main control spool second main control spool

shuttle valve

Pressure relief valve

0 suction valve

2 shuttle valve

4 regeneration valve

8 first pressure regulating valve arrangement 10 second pressure regulating valve arrangement 12 first topping piston

14 second poppet

16, 1 18Ventilbohrung

0 pressure chamber

2, 124 check valve

6 aperture

8, 130Druckraum

32, 134 Pistons

6, 138Blende

piston bore

piston bore Bore ground Bore ground face

face

Axial projection axial projection of the face

face

Individual pressure scale Individual pressure balance Working line Check valve Non-return valve

Claims

claims

1. A hydraulic valve arrangement for supplying pressure medium to a hydraulic consumer (2) having two working ports (A, B) for connection to the consumer (2), with a first inlet orifice (66) via which a first feed stream from the first (A) of Working ports (A, B) towards the consumer (2) is controllable, and a separate second Rücklaufmessblende (90), via which a first return flow from the load (2) via the second working port (B) is controllable, and with a second inlet orifice (68) via which a second feed stream from the second working port (B) towards

Consumer (2) is controllable, and a separate first Rücklaufmessblende (88), via which a second return flow from the consumer (2) via the first working port (A) is controllable, characterized in that two separate pilot valves (44, 46) are provided are and that the first return orifice (88) for controlling the second return flow via the first pilot valve (44) is controllable, and that the second

Inlet metering orifice (68) for controlling the second feed stream simultaneously via the first pilot valve (44) and the second pilot valve (46) is controllable.

2. Valve arrangement according to claim 1, wherein the second return orifice (90) for controlling the first return flow via the second pilot valve (46) is controllable, and wherein the first inlet orifice (66) for controlling the first supply flow simultaneously via the first pilot valve (44). and the second pilot valve (46) is controllable.

3. Valve arrangement according to claim 2, wherein the first pilot valve (44) has a first

Control pressure can be generated, via which the first return orifice (88) and the second inlet orifice (68) are controllable, and wherein the second (46) pilot valve, a second control pressure can be generated via which the second return orifice (90) and the first inlet orifice (66 ) are controllable.

4. Valve arrangement according to claim 3, wherein one of the respectively controlled inlet orifice (66, 68) forming the main spool (92, 94) is acted upon by both control pressures.

5. Valve arrangement according to claim 4 with a selector valve (74) having a first one

Movement direction of the consumer (2) associated with the first position (c), via the one Measuring surface of the first inlet orifice (66) having the first main spool (92) is acted upon by the second control pressure and on a measuring surface of the second inlet orifice (68) having the second main spool (94)

Low pressure or tank pressure can be acted upon, and with a second

Movement direction of the consumer (2) associated second position (b), via which a measuring surface of the second inlet orifice plate (68) having second

Main control slide (94) can be acted upon by the first control pressure and via a measuring surface of the first inlet orifice (66) having first main spool (92) with low pressure or tank pressure can be acted upon.

6. Valve arrangement according to claim 4, wherein the one respective metering orifice forming the main spool (92, 94) with a respective end face (164, 166) each other, and wherein from the respective end face (152, 154) ago a piston bore (140, 142) is introduced, wherein in the respective piston bore (140, 142), a respective pressure piston (92, 94) is guided, each with their end faces (152, 154) in the

Piston bore (140, 142) delimit a control chamber, and with their mutually facing end faces (162, 164) for common axial displacement against each other can be applied, and wherein a respective pressure piston (92, 94) upon displacement in the direction of its associated Main control slide (92, 94) can be applied to this, and wherein in a system of the respective pressure piston (92, 94) on its associated main spool (92, 94) of the brought into abutment of the pressure piston (92, 94) limited control chamber is closed ,

7. Valve arrangement according to the preceding claims, wherein a respective metering orifice (92, 94) an individual pressure compensator (168, 170) is connected downstream.

8. Valve arrangement according to one of claims 1 to 6 with a pressure compensator (24), the

upstream of both inlet flow orifices (66, 68), and via which the inlet orifices (66, 68) can be fluidically connected to the high pressure connection (P), and in

Opening direction with a pressure upstream of the Zulaufmessblenden (66, 68) is acted upon, and in the closing direction with a higher of the pressures on the

Working ports (A, B) or a dependent pressure or a higher of the control pressures is applied.

9. Valve arrangement according to claim 8, with a pressure selector valve (102) via which the higher of the pressures at the working ports (A, B), or the pressure dependent thereon, against the higher of the control pressures can be weighed.

10. Valve arrangement according to one of the preceding claims, with a

Regenerationswegeventil (104), via which the first return flow is at least partially fed to the first feed stream, and / or via which the second return stream is at least partially fed to the second feed stream.

11. Valve arrangement at least according to claim 3 with a control unit in which a characteristic map of the first control pressure in dependence of the second control pressure, and / or in reverse dependence, is stored.

12. Hydraulic valve block having at least one valve arrangement (8), which is designed according to one of the preceding claims, for supplying pressure medium at least one hydraulic consumer (2).

13. Valve block according to, wherein valve slide (92, 94) of the inlet orifices (66, 68) and / or the return flow orifices (88, 90) in a valve bore (118) are arranged together.

14. Valve block with a valve arrangement (8) according to claim 8 and 10, wherein a

Valve spool of the pressure compensator (24) and the Regenerationswegeventils (104) in a valve bore (116) are arranged together.

15. Hydraulic drive with a hydraulic pump (4), at least one hydraulic

Consumer (2) and at least one valve arrangement (8) according to one of claims 1 to 11, via which the at least one hydraulic consumer (2) can be supplied with pressure medium.