EP1934487B1 - Hydraulic control device - Google Patents

Abstract

The device has a measuring orifice supplying a pressurizing medium to a hydraulic load (14). A variable displacement pump (10) is controlled depending on a highest load pressure of actuated hydraulic loads (14, 15) in such a manner that pressure in a supply line (13) lies above the load pressure around a predetermined pressure difference. A load signaling line (38) is connected with a control chamber of a pressure scale (30). A control edge is provided at a valve piston (48) of a control valve (45), with which a flow cross section between the supply line and load signaling line is controlled.

Description

The invention relates to a hydraulic control device for a priority hydraulic consumer and a subordinate hydraulic consumer, according to the preamble of patent claim 1.

Such a hydraulic control circuit is from the DE 197 03 997 A1 known. The two hydraulic consumers while the pressure fluid flows through a metering orifice, wherein the priority assigned to the first priority metering first metering orifice preceded by a pressure compensator and the subordinate second hydraulic consumer associated, second metering orifice downstream of a pressure compensator. Using the pressure compensators constant pressure differences are maintained over the metering orifices with sufficient pressure medium supplied regardless of the load pressures of the hydraulic consumers, so that the hydraulic fluid consumers flowing pressure fluid quantity only depends on Publ tion cross section of the respective metering orifice. The pressure medium source is usually an adjustable hydraulic pump, which is controllable in dependence on the highest load pressure such that the pressure in a supply line to a certain pressure difference above the highest load pressure.

With regard to the first consumer, the control circuit corresponds to a so-called load-sensing or load-sensing control (LS control). From a LS-control or LS consumers are usually referred to when hydraulic consumers are controlled to which pressure medium flows via a metering orifice and an upstream pressure compensator, and when the pressure compensator detects the falling over the respective metering orifice pressure and keeps constant. The pressure compensator is in the closing direction only by Pressure in front of the metering orifice and in the opening direction only by the load pressure of the respective hydraulic consumer and acted upon by a compression spring.

With regard to the second consumer, the control circuit corresponds to a so-called LUDV control. In this case, the second metering orifice downstream pressure compensator is acted upon in the opening direction by the pressure after the second metering orifice and in the closing direction of a pending in a rear control chamber control pressure, which usually corresponds to the highest load pressure of all supplied by the same hydraulic pump hydraulic consumers. If a plurality of hydraulic consumers controlled in this way are actuated at the same time, the quantities of pressure medium flowing to them are proportionally reduced if the quantity of pressure medium supplied by the hydraulic pump is smaller than the required partial quantities of pressure medium. In this case, one speaks of a controller with load-independent flow distribution (LUDV control). The correspondingly controlled hydraulic consumers are called LUDV consumers for short. The LUDV control is a special case of a load-sensing or load-sensing control (LS control). Even with her, the highest load pressure is sensed and generated by the pressure medium source an inlet pressure, which is by a certain amount Ap over the highest load pressure.

The listed publication DE 197 03 997 A1 discloses a priority circuit between the LS consumer and one or more LUDV consumers, in which the LS consumer is primarily supplied with pressure medium. For this purpose, in addition to the pressure compensator of the LS consumer, a priority valve is provided which has a connected to a line section upstream of the first metering orifice first port and a second terminal connected to the Lastmeldeleitung and its valve member in the direction of opening the connection between the first port and the second port from the load pressure of the priority hydraulic consumer, so the LS consumer, and an additional force can be acted upon. In the direction of closing, the valve member is acted upon by the pressure upstream of the metering orifice of the LS consumer - in a supply line or between the pressure compensator and the first metering orifice -. This way becomes a priority Pressure medium supply of the LS consumer ensured. In particular, the pressure upstream of the first metering orifice is controlled to a value that is at least an amount above the load pressure of the primary consumer, which corresponds to the additional force acting on the valve member of the priority valve.

The aim of the invention is based on the described prior art to provide a simpler and cheaper to produce hydraulic control device.

This object is achieved by a hydraulic control device with the features of claim 1.

The present invention provides a hydraulic control device for a priority hydraulic consumer and one or more subordinate consumers. The primary consumer is controlled via a first metering orifice, which is preceded by a (LS) pressure compensator. The subordinate consumer is supplied via a second metering orifice, which is followed by a pressure compensator - in the manner of a LUDV control.

The peculiarity of the present invention is that a further control edge is provided on the valve piston of the pressure compensator of the priority consumer, which controls an injection of pressure medium from a supply line into a load-reporting line. Thus, two control edges are present on the valve piston of this pressure balance. The first control edge controls the pressure medium flow supplied to the first metering orifice in the sense of an individual pressure balance for the primary consumer. The second control edge controls a flow cross-section between the inlet and the load-sensing line. As a result, the pressure in the load-sensing line can be increased if the pressure difference dropping at the metering orifice of the 'primary consumer drops below a certain value. This causes an increase in the pressure level downstream of the second metering orifice and thus a reduction in the pressure medium flow supplied to the subordinate consumers. As a result, pressure medium is available in sufficient quantity to the primary consumer.

The present invention takes advantage of the recognition that the pressure balance of the primary consumer and the control mechanism of a pressure increase in the load reporting line are controlled by the same pressure signals skillful to realize these two functionalities in a single, simple design valve. Compared to the conventional solution, a separate priority valve and thus material, space and cost is saved. Due to the smaller number of movable components, the control device according to the invention is also low maintenance.

Furthermore, the pressure difference arising at the metering orifice of the primary consumer can be kept virtually constant, regardless of the operating state, since this pressure difference is determined in each operating state by the control spring of the pressure compensator. With sufficient flow rate of the pump and load-carrying subordinate consumers, the pressure compensator of the priority consumer behaves like an individual pressure compensator and throttles the pressure medium supply so that a specific pressure difference prevails over the metering orifice of the primary consumer. If there is a supersaturation, the pressure in the load-sensing line is regulated by the second control edge, so that in turn the pressure difference at the metering orifice of the primary consumer corresponds to the pressure equivalence of this control spring. In comparison, in conventional control, there are various springs in the pressure compensator and in a priority valve which controls the load-sensing line. In order to ensure a clearly determinable system behavior in view of finishing tolerances, these springs are set to different pressure equivalence values. Thus, in the conventional system, there may be a noticeable pressure reduction at the metering orifice of the primary consumer e.g. at the transition to supersaturation.

Further advantageous embodiments are specified in the subclaims.

Thus, according to claim 2, the control edges are arranged so that a direction of movement for opening the first flow cross section corresponds to the direction of movement for opening the second flow cross section. This means that the control edges are formed on axially aligned in the same direction surfaces of the valve piston. The same direction actuation of the two control mechanisms of the pressure compensator simplifies their implementation by means of a single valve piston.

A particularly preferred embodiment provides that an opening of the second Durchflußquefschnitts - ie a pressure medium supply to the load-sensing line - only then takes place when the hydraulic resistance at the first flow area is almost minimal. This means that the control mechanism of an increase in pressure in the load-sensing line only starts when the regulation of the flow rate through the pressure compensator - more precisely through the first flow cross-section - has reached the upper flow limit of its control range. As a result, an unnecessary increase in the pressure level of the variable displacement pump and throttling of the subordinate LUDV consumers is avoided as long as the variable displacement pump still supplies sufficient pressure medium. If the control ranges of these two control mechanisms connect to each other and not or only slightly overlap, also always a stable, the respective load conditions clearly assigned operating state of the hydraulic control device according to the invention is always guaranteed.

A simple design of the pressure balance of the priority consumer results when it is designed as a slide valve with a valve bore and inlet chamber formed thereon and with two Ausgangskammem - a first connected to the metering orifice output chamber and a second connected to the Lastmeldeleitung output chamber.

The complexity of the pressure balance of the primary consumer can be reduced if one end face of the valve piston to the first output chamber, which is connected to the metering orifice, adjacent. As a result, the pressure in the output chamber simultaneously acts as control pressure on the valve piston in order to pressurize it in the closing direction of the two flow cross sections.

According to a further preferred embodiment, a fluid path is formed in the valve piston, which connects a control pressure chamber formed on an end face of the valve piston with the first output chamber. Such a fluid path is easy to manufacture and represents a space-saving way to pressurize a control pressure chamber of the pressure compensator with the pressure upstream of the metering orifice.

Preferably, the fluid path comprises a bore which opens into the peripheral surface of the valve piston and which can be brought into overlap with the second output chamber. This achieves a particularly advantageous construction of the pressure balance, since the fluid path also serves as a pressure line to the control pressure chamber and as a flow path into the second output chamber. In any case, only a small flow cross-section is required between the supply line and the load-signaling line, and a quantity of pressure medium supplied / discharged to the control pressure chamber is likewise small. Thus, a small diameter fluid path can be used. Since a fluid supply to the load-sensing line should only take place when the Dürchflussquerschnitt between the inlet chamber and the first output chamber of the pressure compensator is already largely open, the fluid pressure in the first output chamber largely corresponds to the fluid pressure in the inlet chamber. Therefore, fluid from a portion of the first output chamber can be easily supplied to the second output chamber instead of directly from the inlet chamber, and the described simple valve construction can be achieved.

Alternatively, a depression is present in the valve piston, which can be brought into coincidence with the inlet chamber and the second outlet chamber at the same time. As a result, if necessary, the first flow cross section and the second flow cross section can be designed independently of each other.

Hereinafter, the present invention and its advantages will be explained in more detail with reference to the embodiment shown in the figures.

Fig. 1

ein Schaltbild einer hydraulischen Steuervorrichtung mit einem vor- rangigen Verbraucher und einem Regelventil, das den Druckmittelstrom zum vorrangigen Verbraucher sowie zusätzlich eine Druckmitteleinspeisung in eine Lastmeldeleitung steuert.

Fig. 2

ein Schnittbild des in Fig. 1 dargestellten Regelventils,

Fig. 3A

ein Schnittbild des in Fig. 1 dargestellten Regelventils in einer alterna- tiven Ausführung,

Fig. 3B

eine Symboldarstellung des in Fig. 3A gezeigten Regelventils,

Fig. 4

eine Prinzipskizze einer weiteren Ausführung des in Fig. 1 dargestell- ten Regelventils und

Fig. 5

ein Schaltbild einer hydraulischen Steuervorrichtung, vergleichbar der Figur 1, mit einer Bypassleitung zum Melden eines Lastdrucks des vorrangigen Verbrauchers in die Lastmeldeleitung.

Show it:

Fig. 1

a circuit diagram of a hydraulic control device with a priority consumer and a control valve, the Pressure medium flow to the priority consumer and additionally controls a pressure medium feed into a load-sensing line.

Fig. 2

a sectional view of the in Fig. 1 illustrated control valve,

Fig. 3A

a sectional view of the in Fig. 1 illustrated control valve in an alternative embodiment,

Fig. 3B

a symbol representation of the in Fig. 3A shown control valve,

Fig. 4

a schematic diagram of another embodiment of the in Fig. 1 illustrated control valve and

Fig. 5

a diagram of a hydraulic control device, comparable to FIG. 1 , with a bypass line for reporting a load pressure of the priority consumer in the load reporting line.

To Fig. 1 sucks a variable 10 with an adjustment. 11 pressure fluid from a tank 12 and it is in a system of supply lines. Via the supply lines, a first hydraulic consumer 14, which is designed as a synchronous cylinder, and at least a second hydraulic consumer 15, which is a differential cylinder, supplied with pressure medium. The direction and velocity of the synchronizing cylinder 14 is determined by a corresponding operation of a 4/3-way proportional valve 16 whose valve spool is spring-centered in a mid-position in which the four working ports and a control port 18 of the directional control valve 16 are locked. In a displacement of the valve spool from its center position out in one or the other direction, a metering orifice 17 is opened differently wide depending on the way the valve spool is moved. The control port 18 is connected downstream of the metering orifice with the flow to the synchronizing cylinder 14.

Between a supply line 13 and an inlet port 19 of the directional control valve 16, a control valve 45 is inserted, which has the function of a 2-way pressure compensator. Accordingly, the control valve 45 controls the flow area of a fluidic connection 20 between its inlet 46 and one of its outlets 23, ie between the inlet 13 and the inlet 19 Directional valve 16. The valve piston 48 of the control valve 45 is in the direction of closing the connection 20 from the pressure upstream of a metering orifice 17 and in the direction of opening via a control line 61 from the pressure in the control port 18 of the directional control valve 16, ie the load pressure of the Gleichgangzylinders 14, and of a control spring 21 charged. The force of the control spring 21 is designed so that its a pressure difference of z. B. 15 bar above the metering orifice 17 is equivalent.

While thus the first hydraulic consumer 14 associated control valve 45 of the first metering orifice 17 is connected upstream, the second hydraulic consumer 15 associated second pressure compensator 30 of a second metering orifice 31 downstream, for the directional control of the differential cylinder 15 is between the second pressure compensator 30 and the differential cylinder a directional control valve 32 is arranged, over which compared to the pressure drop at the metering orifice 31, no appreciable pressure drop occurs more when the differential cylinder 15 is actuated. The metering orifice 31 and the control grooves necessary for directional control are formed in a known manner on the same valve spool, so that direction and speed control in each case go readily together. The control piston 33 of the pressure compensator 30 is in the direction of opening the connection between the metering orifice 31 and the directional control valve 32 from the pressure to the metering orifice and in the direction of closing the connection of a prevailing in a rear control pressure chamber 34 control pressure and a weak compression spring 35, the pressure of z: B. only 0.5 bar equivalent is applied. The front side of the control piston 33 is connected to the control pressure chamber 34 via a channel 36 extending in the control piston, wherein a check valve 37 opening towards the control pressure chamber is arranged in the channel 36.

Parallel to the metering orifice 31, the pressure compensator 30 and the directional control valve 32 for the second hydraulic consumer 15, additional metering orifices, pressure compensators and directional control valves for further hydraulic consumers can be connected to the system of the supply lines 13. The control pressure chambers 34 of all pressure compensators 30 are connected to each other, so that in this Control pressure chambers the same pressure is pending. The control piston 33 of the pressure compensators are trying to bring in an operation of a second hydraulic consumer in such a position in which sets at its front side only by the force of the compression spring 35 equivalent pressure difference higher pressure than in the control pressure chambers 34th

Via the channels 36 and the check valves 37, once the first hydraulic consumer 14 is completely disregarded, the highest load pressure of all actuated second hydraulic consumers 15 is let into the control pressure chambers 34. The control pressure chambers 34 are connected to a load signaling line 38, which leads to the adjustment 11 of the pump 10. The load-signaling line 38 is furthermore connected to the tank 12 via a current regulator 55. About this current regulator the load-sensing line 38 is relieved of pressure, if none of the hydraulic load is operated.

Variable pumps and corresponding regulators are well known and readily available on the market. It is therefore unnecessary to go into detail. It should merely be pointed out that the pump regulation causes a pressure to be set in the supply line 13 which is above the pressure in the load-reporting line 38 by a pressure difference Δp equivalent to the force of a control spring. The pressure difference .DELTA.p amount z. B. 20 bar, so is higher than the force of the control spring 21 of the control valve 45 equivalent pressure difference of 15 bar.

The first hydraulic consumer 14 is to be supplied with priority prior to the second hydraulic consumer 15 with pressure medium. For this purpose, a second controllable connection 22 is provided in the control valve 45. The connection 22 is formed as a diaphragm with proportionally controllable Durchftussquerschnitt between the input 46 and an output 47. The latter is connected to the load-signaling line 38.

Also for the second fluidic connection 22 controlled by it, the valve piston 48 of the control valve 45 is in the closing direction of a pressure upstream of the metering orifice 17 and in the direction of opening of the through Control line 61 zoom supplied load pressure of the primary consumer 14 and the control spring 21 acted upon.

In the FIG. 2 the control valve 45 is shown in more detail. In the valve housing 70, a valve bore 71 is present. In this bore, the valve piston 48 is slidably mounted. An inlet chamber 72 and two outlet chambers 73 and 74 adjoin the valve bore. The inlet chamber is connected to the port 46 designed as a bore and thus to the supply line 13. The outlet chamber 73 communicates with the outlet 23, ie with the metering orifice 17. The output chamber 74 opens via the connection 47 in the load-signaling line 38.

The controllable fluidic connection 20 is achieved via a radially recessed portion 76 of the valve piston 48. At a stage arranged in the direction of the inlet chamber 72, the control edge 77 is formed on the valve piston 48. The control edge 77 defines a first flow area between it and a housing web 78 'formed between the inlet chamber 72 and the outlet chamber 73.

The fluidic connection 22 is formed by a recess 78 in the peripheral surface of the valve piston 48. This recess 78 may be e.g. be an axial groove or a radial graduation of the valve piston. A control edge 79 delimiting the depression 78 in the direction of the outlet chamber 74 forms a second controllable and closable flow cross-section with the outlet chamber 74.

wobei p_Ls der Lastdruck des vorrangigen Verbrauchers 14, p₂₁ das Druckäquivalent der Kraft der Regelfeder 21, p₃₈ der in der Lastrneldeleitung 38 herrschende Lastdruck, Δp die Regeldruckdifferenz der Pumpenverstellung 11 und Δp_DW der an der Steuerkante 77 der Verbindung 22 abfallende Druck ist.A control pressure chamber 50 is connected to the control line 61, which carries the load pressure of the priority consumer 14. The pressure in the control pressure chamber 50 acts on the valve piston 48 in the opening direction of the fluid connections 20 and 22. In addition, the force of the control spring 21 on the valve piston 48 in the opening direction. In the closing direction, the pressure prevailing in the control pressure chamber 49 acts. The control pressure chamber 49 is connected via a formed in the valve piston 48 fluid passage 75 with the radially recessed portion 76 and thus the output chamber 73 in fluid communication. Based on FIGS. 1 and 2 Now, the operation of the control device according to the invention will be explained. At the valve piston 48 of the control valve 45, an equilibrium of forces arises that takes account of the following force or pressure components: $${\mathrm{p}}_{\mathrm{LS}}\mathrm{+}{\mathrm{<figure-callout\; id="21"\; label="p"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{21}}\mathrm{=}{\mathrm{<figure-callout\; id="38"\; label="p"\; filenames="imgf0001.png,imgf0005.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{38}}\mathrm{+}\mathrm{Ap}\mathrm{-}{\mathrm{Ap}}_{\mathrm{DW}}$$

where p _{Ls is} the load pressure of the primary consumer 14, p _{21 is} the pressure equivalent of the force of the control spring 21, p _{38 is} the load pressure prevailing in the service line 38, Δp is the control pressure difference of the pump displacement 11 and Δp _{DW is} the pressure drop at the control edge 77 of the connection 22 ,

The control edge 79 is arranged so that the connection 22 opens only when the flow cross section at the control edge 77 is already almost maximum, ie when the pressure drop Δp _DW at the control edge 77 has a value Δp _DW . has reached close to a minimum. The value Δp _DW . However, it depends on the flow rate at the control edge 77.

If the pressure medium flow rate conveyed by the pump is sufficient to supply all the consumers, then the control pressure difference Δp remains constant at the value set by the control spring of the pump adjustment 11, e.g. 20 bar.

As long as the secondary consumers are load-guiding, more specifically, the pending in the supply line 13 pressure sozulange greater than the sum of the load pressure of the priority consumer 14 and the pressure equivalence of the control spring 21 is, via control edge 77, a pressure drop Ap _DW for controlling the supply of the primary consumer. Due to the pressure drop Δp _DW , a pressure surplus existing from the point of view of the primary consumer 14 in the supply line 13 is throttled. The pressure p ₃₈ in the load-sensing line corresponds to the highest load pressure of the subordinate consumers, hereinafter referred to as p _LUDV .

wobei (p_LUDV + Δp) der durch die nachrangigen Verbraucher 15 erzeugbare Zulaufdruck ist, ist der Steuermechanismus einer Drosselung an der ersten Steuerkante 77 erschöpft und der zugehöhrige Durchflussquerschnitt ist vollständig geöffnet.At a load pressure of the priority consumer of $${\mathrm{p}}_{\mathrm{LS}}\mathrm{>}\left({\mathrm{p}}_{\mathrm{LUDV}\mathrm{+}\mathrm{Ap}}\right)\mathrm{-}{\mathrm{p}}_{\mathrm{21}}\mathrm{-}{\mathrm{\Delta p}}_{\mathrm{DW}*}$$

where (p _LUDV + Δp) is the inlet pressure that can be generated by the subordinate consumer 15, the control mechanism is throttling at the first one Control edge 77 is exhausted and the associated flow cross-section is fully open.

wenn man berücksichtigt, dass die Regelfeder 21 beim Regeln an der Steuerkante 79 etwas stärker entspannt ist als beim Regeln an der Steuerkante 77, also eine etwas geringere Druckäquivalenz p₂₁' als p₂, aufweist, und wenn Δp_DW* als ein geringer Druckabfall an dem von der Steuerkante 77 begrenzten nahezu vollständig geöffneten Durchflussquerschnitt angenommen wird. Im Wesentlichen regelt sich also der Druck in der Zuleitung 13 auf einen Wert ein, der um die Druckäquivalenz p₂₁' der Regelfeder 21 höher als der Lastdruck des vorrangigen Verbrauchers 14 ist.So if the inlet pressure (p ₃₈ + Δp) is at or below the value p _LS + p ₂₁ + Δp _DW . decreases, the second control edge 79 opens the flow cross-section of the connection 22. This increases the pressure p ₃₈ in the load-sensing line 38 to values greater than p _LUDV . If the pressure in the supply line 13 (p ₃₈ + Δp) was _previously dependent only on the load pressure p _{LUDV of} the subordinate consumers, the supply pressure (p ₃₈ + Δp) is now determined by the load pressure p _{LS of} the priority consumer 14. The regulation of the supply line pressure (p ₃₈ + Δp) takes place with the aid of the control edge 79 and the feedback via the adjustment 11. Equation 1 results directly in the dependency $$\left({\mathrm{p}}_{}\right)$$$$\left({\mathrm{}}_{\mathrm{38}}\mathrm{+}\mathrm{Ap}\right)\mathrm{=}{\mathrm{p}}_{\mathrm{LS}}\mathrm{+}{\mathrm{<figure-callout\; id="21"\; label="p"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">p</figure-callout>}}_{\mathrm{21}}\mathrm{\text{'}}\mathrm{+}{\mathrm{Ap}}_{\mathrm{DW}*}$$

when one considers that the control spring 21 when controlling at the control edge 79 is slightly more relaxed than when controlling at the control edge 77, so a slightly lower pressure equivalence p ₂₁ 'than p ₂ , and if Δp _DW * as a small pressure drop at assumed by the control edge 77 almost completely open flow cross-section is assumed. In essence, therefore, the pressure in the supply line 13 is adjusted to a value which is higher than the load pressure of the priority consumer 14 by the pressure equivalence p ₂₁ 'of the control spring 21.

In the event that the pressure medium flow delivered by the pump 10 is not sufficient to supply all consumers, Δp can no longer be regarded as constant. The control capacity of pump 10 and its adjustment 11 are exhausted and the pressure in the supply line 13 decreases. As before, the connection 22 opens when the supply pressure drops (p ₃₈ + Δp) to p _LS + p ₂₁ + Δp _{DW *.} This leads to an increase in the pressure conducted in the load-signaling line 38. This also increases the pressure between the metering orifice 31 and the pressure compensator 30 of the subordinate consumer. The above the metering orifice 31 pending pressure difference is reduced and thus also the subordinate consumer feedable pressure medium flow. If necessary, with correspondingly reduced control pressure difference Δp, the pressure of the load-sensing line 38 increase to the inlet pressure (p ₃₈ + Δp) and completely suppress the supply of the subordinate consumer 15 by means of the pressure compensator 30. Also, several subordinate consumers 15 can be abgeregelt in this way. By means of this mechanism of throttling of the subordinate consumers 15, the inlet pressure (p ₃₈ + Δp) in turn adjusts according to Equation 3 to a value which is substantially higher than the load pressure of the primary consumer 14 by the pressure equivalence p ₂₁ 'of the control spring 21.

In all cases, therefore, a reliable supply of the priority consumer 14 is secured so that one of the pressure equivalence p ₂₁ and p ₂₁ 'of the control spring 21 corresponding pressure difference across the metering orifice 17 is applied.

The FIG. 3A shows a control valve 85, which represents a modified embodiment of the control valve 45. A symbol representation of the control valve 85 is in FIG. 3B shown. The control valve 85 differs only by the valve piston 88 of the control valve 45. Like the valve piston 48, the valve piston 88 has a radially recessed piston portion 76. From this piston portion is a fluid channel 75 goes out, which opens at an end face of the valve piston 88 in the control pressure chamber 49 , In contrast to the valve piston 48, there is no recess in the piston circumferential surface with which the inlet chamber 72 can be connected directly to the outlet chamber 74. Instead, a bore 86 is made perpendicular to the axis of the valve piston 88, which opens into the fluid channel 75. Together with a Feinsteuemut 87, the bore 86 forms a control edge 89 for controlling a flow cross section at the output chamber 74. It should be noted that this formed between the control edge 89 and the valve housing 70 opening cross-section opens only when the hydraulic resistance or the pressure drop .DELTA.p _DW at the control edge 77 has already reached a value Δp _{DW *} near the minimum value. Thus, the pressure of the pressure medium, which is supplied via the fluid path 75 when opening the control edge 89 of the radially recessed piston portion 76 corresponds approximately to the pressure in the input port 46. Thus, the pressure in the Lastmeldeleitung 38 bis be increased almost to the pending at the input port 46 supply pressure.

Another embodiment of a control valve 95, which may take the place of the control valve 45 or 85 is in FIG. 4 outlined. The symbol representation of the control valve 95 corresponds to the FIG. 3B , In the valve housing 90 of the control valve 95, a valve bore 91 is provided. On the valve bore 91, an inlet chamber 92 and two outlet chambers 93 and 94 are arranged. The chambers 92, 93 and 94 are fluidly connected to respective ports 46, 47 and 23, as in FIGS FIG. 4 shown. In the valve bore 91, a cylindrical valve piston 96 is movably guided. The valve piston 96 has an axially extending blind hole 97 open in the direction of the outlet chamber 93. From the peripheral surface of the valve piston 96, two radially extending holes 98 and 99 abut the blind hole 97.

The bore 98 can be brought into overlap with the inlet chamber 92. As a result, a fluid connection is created by the inlet connection 46 via the bore 98, the blind bore 97, the outlet chamber 93 and the outlet connection 23. The control edge 100, which controls the flow cross-section of this connection, is the circumferential surface-side edge of the bore 98. A fluid connection from the inlet connection 46 to the outlet connection 47 is established via the bore 98, the blind bore 97, the bore 99 and the outlet chamber 94. The relevant control edge 101 is the circumferential surface side edge of the bore 99. The bore 99 is arranged so that it only overlaps with the output chamber 94, when the controlled at the bore 99 flow cross-section only a small hydraulic resistance / pressure drop Δp _{DW *} caused. As a result, the pressure in the load-signaling line 38 can be increased to almost the inlet pressure at the inlet connection 46.

At one of the blind hole 97 facing away from the end face of the valve piston 96 limits this formed in the valve housing 90 control pressure chamber 50. This is connected to the control line 61, the Load pressure of the priority consumer 14 leads. The pressure prevailing in the control pressure chamber 50 acts in the opening direction of the controlled through the holes 98 and 99 compounds. In addition, the control spring 21 arranged in the control pressure chamber 50 acts in the opening direction. In the closing direction of the valve piston 96 is acted upon directly by the pending in the output chamber 93 pressure, since the valve piston 96 with its end face in which the blind hole 97 opens, adjacent to the output chamber 93. On the one hand by this embodiment of the control valve 95, a very low pressure drop Δp _{DW *} at the bore 98 achievable, on the other omitted by the arrangement of the output chamber 93 at the front end of the valve piston 96, a separately formed control chamber and an internally or externally guided to this control line ,

The FIG. 5 shows a further embodiment of the hydraulic control device according to the invention. The execution after Fig. 5 is largely the same after execution Fig. 1 , Different from the execution after Fig. 1 is that now the control line 61, which leads from the control port 18 of the directional control valve 16 to the control valve 45, via a located in a bypass line 62 check valve 63 is also connected to the load signaling line 38. In this case, the check valve 63 locks from the load-signaling line 38 to the channel 61, that is, to the control port 18 of the directional control valve 16. Furthermore, a check valve 64 is also arranged between the second connection 47 of the control valve 45 and the load-receiving line 38. This locks to port 47.

In the execution after Fig. 1 finds a change in the control mechanism of the control valve 45 takes place even if in sufficient quantity conveyed pressure medium, as described above, if the load pressure p _{LS of} the priority consumer 14 by the subordinate consumer predetermined inlet pressure (p _LUDV + Δp) minus the pressure equivalent p ₂₁ of Control spring 21 exceeds (the pressure drop Δp _{DW *} at the control edge 77 is negligible). If the primary consumer becomes load-bearing in this sense, the control valve loses its functionality as a LS pressure compensator. This is replaced by the mechanism of controlling the pressure in the load-sensing line 38.

In the execution after Fig. 5 is carried out at sufficiently promoted pressure fluid quantity and load-carrying priority hydraulic consumers 14 of the load pressure gravel hydraulic consumer via the check valve 63 in the load reporting line 38. The pressure in the supply line 13 is thus the Regeleldruckdiflerenz .DELTA.p of the variable displacement pump 10 above the load pressure of the hydraulic consumer 14. The control valve 45 has in this case the function of a LS-pressure compensator and throttles the metering orifice 17 supplied pressure medium flow through the first control edge 77th Thus, the pressure difference across the pressure compensator 17 corresponds to the pressure equivalent p _{21 of} the control spring 21.

Only when the saturation pressure (p ₃₈ + Δp) in the supply line 13 to the sum of the load pressure p _{LS of} the hydraulic consumer 14, the pressure equivalent p _{21 of} the control spring 21 and a small pressure drop Δp _{DW *} at the control edge 77 has dropped the load-sensing line 38 is supplied via the connection 22 pressure medium. A reduction of the pressure drop across the metering orifice 17 virtually does not occur, because with further increasing undersaturation, the pressure p ₃₈ in the load-reporting line 38 is increased via the control valve 45 and thereby the pressure compensators 30 of the LUDV consumers 15 are adjusted in the closing direction.

The check valve 64 prevents a pressure medium flow from the hydraulic consumer 14 via the check valve 63 in the system of the supply lines, if z. B. at the beginning of an actuation of the pressure in the supply lines is not above the load pressure.

The check valve 64 may also be omitted if the port 47 of the control valve 45 is connected to the check valve 63 so that the check valve 63 to port 47 blocks out.

LIST OF REFERENCE NUMBERS

10

variable

11

adjustment

12

tank

13

supply

14

Rod cylinders

15

differential cylinder

16

4/3-way proportional directional valve

17

metering orifice

18

control connection

19

inflow connection

20

Fluidic connection

21

control spring

22

Fluidic connection

23

output

30

pressure compensator

31

metering orifice

32

way valve

33

control piston

34

Control pressure chamber

35

compression spring

36

channel

37

check valve

38

Load-sensing line

45

control valve

46

entrance

47

output

48

plunger

49

Control pressure chamber

50

Control pressure chamber

55

current regulator

61

control line

62

Bypass line

63

check valve

64

check valve

70

valve housing

71

valve bore

72

inlet chamber

73

exit chamber

74

exit chamber

75

fluid channel

76

Reverse piston section

77

control edge

78

deepening

78 '

housing web

79

control edge

85

control valve

86

drilling

87

Feinsteuernut

88

plunger

89

control edge

90

casing

91

valve bore

92

inlet chamber

93

exit chamber

94

exit chamber

95

control valve

96

plunger

97

Blind hole

98

radial bore

99

radial bore

100

control edge

101

control edge

Claims (8)

1. A hydraulic control device for a primary-priority, first hydraulic consumer (14) and for a secondary-priority, second hydraulic consumer (15)
   having a first metering orifice (17), via which pressure medium can be conducted to the first hydraulic consumer (14),
   having, upstream of the first metering orifice (17), a pressure compensator (45) by means of which a constant pressure differential can be adjusted across the first metering orifice (17) and which is provided for this purpose with a valve piston (48) with a first control edge (77) with which a first flow cross-section between an inlet line (13) and the first metering orifice (17) can be controlled,
   having a second metering orifice (31) via which pressure medium can be conducted to the second hydraulic consumer (15) and to which is connected on the downstream side a second pressure compensator (30) which in the closing direction can be acted upon by a control pressure prevailing in a control chamber (34) and in the opening direction by the pressure after the second metering orifice (31),
   having a pressure medium source (10) with variable delivery flow which can be controlled as a function of the highest load pressure of the actuated hydraulic consumers (14, 15) in such a manner that the pressure in the inlet line (13) lies above the highest load pressure by a certain pressure difference, and
   having a load sensing line (38) which can be acted upon by the load pressure of the second hydraulic consumer (15) or by a pressure derived therefrom and which is connected to the control chamber (34) of the second pressure compensator (30) as well as to a control element (11) of the pressure medium source (10),
   characterized by the fact that
   on the valve piston (48) of the first pressure compensator (45) there is provided a second control edge (79) with which a second flow cross section between the inlet line (13) and the load sensing line (38) can be controlled.
2. A hydraulic control device according to claim 1, characterized by the fact that the first control edge (77) and the second control edge (79) are arranged such that a movement direction of the valve piston (48) for opening the first flow cross-section corresponds to a movement direction of the valve piston (48) for opening the second flow cross-section.
3. A hydraulic control device according to claim 1 or 2, characterized by the fact that an opening of the second flow cross-section occurs only when the hydraulic resistance at the first flow cross-section is almost minimal.
4. A hydraulic control device according to any of claims 1 to 3, characterized by the fact that in a valve housing (70) of the pressure compensator (45) there is formed a valve bore (71) in which the valve piston (48) is movably guided and that at the valve bore (71) there are arranged an inlet chamber (72), a first outlet chamber (73) which is fluidically connected to the first metering orifice (17) and a second outlet chamber (74) which is fluidically connected to the load sensing line (38).
5. A hydraulic control device according to claim 4, characterized by the fact that the valve piston (96) is, at an end face, contiguous to the first outlet chamber (93).
6. A hydraulic control device according to claim 4, characterized by the fact that in the valve piston (48; 88) there is formed a fluid path (75) which connects a control pressure chamber (49) contiguous to an end face of the valve piston (48; 88) to the first outlet chamber (73).
7. A hydraulic control device according to claim 6, characterized by the fact that the fluid path (75) has a bore (86) debouching in the circumferential surface of the valve piston (88) which can be brought to overlap with the second outlet chamber (74).
8. A hydraulic control device according to any of claims 4 to 6, characterized by the fact that the valve piston (48) has a depression (78) which can be brought to overlap simultaneously with the inlet chamber (72) and the second outlet chamber (74).

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