EP1629209B1 - Hydraulic control arrangement - Google Patents

Abstract

A hydraulic control arrangement is disclosed for the load-independent control of a consumer with a continuously adjustable distribution valve having a pressure compensator down the line. According to the invention, the pressure compensator has a single-sided damping such that the movement in the opening direction is damped and the movement in the closing direction is substantially undamped. Furthermore, a control arrangement is disclosed in which a load-holding function is integrated in the pressure compensator.

Description

The invention relates to a hydraulic control arrangement for load-independent control of a consumer according to the preamble of claim 1 and a pressure compensator for such a control arrangement.

The basic structure of such control arrangements is known, for example, from WO 95/32364 A1. In such a LUDV system, each consumer is assigned an adjustable metering orifice with a downstream pressure compensator, the latter keeping the pressure drop across the metering orifice constant, so that the quantity of fluid flowing to the corresponding hydraulic consumer depends on the opening cross section of the metering orifice and not on the load pressure of the consumer or on the pump pressure depends. Since, for example, in mobile implements a plurality of such valve assemblies are connected in parallel, is achieved by the individual pressure compensators of the system that in the case where a hydraulic pump of the system has been adjusted to the maximum displacement and the pressure medium flow is insufficient to the predetermined pressure drop across To maintain the metering orifices of each associated with a consumer valve assemblies, the pressure compensators of all actuated hydraulic consumers are adjusted in the closing direction, so that all pressure medium flows are reduced by the same percentage. Because of this load-pressure-independent flow distribution (LUDV), all actuated consumers then move at a speed that is reduced by the same amount.

Such LUDV hydraulic systems are being used to an ever-increasing extent in mobile implements with combined movements. The working movements of these mobile working tools (mini and compact excavators, backhoe loaders, telescopic loaders, skid steer loaders, etc.) should follow the control of the driver without vibration and pressure. It has been proven that damping of the LUDV pressure compensators is required for vibration-free control.

An attenuation is known, for example, from US Pat. No. 6,532,989 B1. In this known solution, the pressure compensator has a rear pressure chamber and an annular damping chamber, both of which can be acted upon by a pressure acting on a pressure compensator piston in the closing direction pressure, while on a front end face of the pressure compensator piston of the downstream of the metering orifice pressure applied, usually the load pressure of the driven Consumer, acts in the opening direction. Between the rear pressure chamber and the damping chamber, a damping nozzle is provided, through which the pressure medium must flow during the axial displacement of the pressure compensator piston out of the damping chamber or into this, so that the pressure compensator piston movement is damped. Such damping inevitably leads to delays in opening and closing the pressure compensator, resulting in a delayed start of movement of high-load work movements.

In contrast, the present invention seeks to provide a control arrangement and a suitable LUDV pressure compensator, in which the delay of the working movement of a consumer is minimized despite the damping of the pressure compensator.

This object is achieved with regard to the hydraulic control arrangement by the features of claim 1 and with respect to the pressure balance by the features of the independent claim 12.

According to the invention, in addition to the damping chamber connecting the damping chamber with the pressure chamber, a connecting recess with a larger cross section is provided, via which the damping chamber is connected to a rear pressure chamber, which can be shut off by a non-return valve opening in the direction of the damping chamber. By this measure, the movement of the pressure compensator piston in the opening direction is relatively strongly damped in dependence on the diaphragm cross section, while in the closing direction opens the check valve and thus aufsteuert a comparatively large cross-section -. the pressure compensator is damped on one side, so that the pressure balance of a consumer with low load pressure, for example, quickly closes and in this way allows the rapid pressure build-up to a higher load pressure in another disc.

In a preferred embodiment, the pressure balance piston is designed as a stepped hollow piston, as described in US 6,532,989 B1. This hollow piston is guided on an axial inner part, which is provided with a blind hole, which opens into the rear pressure chamber. A formed by the stepped portion of the pressure compensator piston inner ring end surface bounded with a correspondingly formed portion of the inner part of the damping chamber. The foot-side annular end face of the stepped piston is acted upon in the opening direction of the pressure compensator by the pressure downstream of the metering orifice.

In the known solutions, a rear control chamber of the pressure compensators is connected to the load reporting line, in which abuts the tapped via a shuttle valve chain highest load pressure of all driven consumers. Now, if the load pressure of an actuated hydraulic consumer quickly rises above the currently prevailing maximum load pressure, so immediately increases the pressure at the front of the pressure compensator piston of the associated pressure compensator, while a corresponding increase in pressure in the rear control chamber via the shuttle valve chain and the Lastmeldeleitung takes place delayed. The consequent short-term power imbalance on the control piston of the pressure compensator can adversely affect the control of the hydraulic load. For example, the hydraulic load can briefly sag slightly or the load-independent flow distribution can be disturbed.

To avoid such unwanted sagging of the consumer additional load holding valves are connected in the pressure medium flow path between the consumer and the pressure compensator in the aforementioned solutions, so that the outflow of the pressure medium can be prevented by the consumer via the pressure compensator. However, such additional load-holding valves make the control arrangement more expensive and require considerable space.

To overcome this disadvantage, pressure compensators are proposed in US Pat. No. 5,067,389, US Pat. No. 5,890,362 and US Pat. No. 4,787,294, in which the load position is integrated into the function of the pressure compensator. In this case, the pressure compensator is provided with two successively arranged pressure compensator pistons which are connected in such a way that the pressure compensator is closed when the pressure applied to the inlet of the pressure compensator is open Pressure compensator piston is lower than the individual load pressure.

In DE 40 05 966 C2, a solution is proposed in which in the pressure compensator piston a shuttle valve is integrated, via which the pressure downstream of the metering orifice and in the load reporting channel is compared and reported in the rear control room.

In DE 296 17 735 U1 a pressure compensator is described, in which the load is reported on a complicated shuttle valve circuit with check valves and nozzles to keep the pressure balance of the load-holding function in the closed state.

All the above-described known solutions with a load-holding function in the pressure balance have the disadvantage that a considerable effort is required to tap a control pressure, which acts on the pressure compensator piston in the load holding function in the closing direction.

According to an embodiment, which can also be claimed independently of claim 1, the damping chamber is connected via the damping nozzle to the channel carrying the individual load pressure, so that in the case where the pressure at the inlet of the pressure compensator drops below this load pressure, the pressure compensator piston moves beyond the pressure compensator piston Damping space applied individual load pressure is brought into its closed position, so that the pressure balance with the load holding function with takes. Compared with the previously described solutions with load-holding function, the construction according to the invention is characterized by an extremely compact and simple construction.

Alternatively, the damping nozzle also connect the damping chamber with the rear pressure chamber, although the load-holding function is dispensed with.

It is preferred if a transverse bore opening in the blind hole is provided at a foot-side end portion of the inner part, which is completely opened in the open position of the pressure compensator piston, so that the pressure is tapped downstream of the metering orifice and guided into the rear pressure chamber.

In a particularly preferred embodiment, a bore or a recess is formed at the smaller diameter of the pressure compensator piston, which can be positioned with respect to the transverse bore, that the pressure downstream of the metering orifice is reported in the blind hole.

In an alternative solution according to the invention, this connection between the channel downstream of the metering orifice and the rear pressure chamber is always open. In a preferred solution, however, this compound is opened only during the initial stroke (seen from the closed position) of the pressure compensator and fully open pressure compensator, while in the intermediate region of this connection is closed, so that the rear pressure chamber is then subjected to the highest effective load pressure while at the beginning of the up-steer of the pressure compensator with the pressure downstream of the orifice plate - ie about the pump pressure - is applied.

The check valve according to the invention may be formed by a simple O-ring, which is placed on the inner part or by a in one Closed position pre-tensioned closing disc. Alternatively, conventional check valves can be used with spring-biased closing bodies.

The pressure compensator piston can be biased by a comparatively weak control spring in the closed position.

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

• Figur 1 eine Schnittdarstellung einer Ventilscheibe mit einer halbseitig gedämpften LUDV-Druckwaage;
• Figur 2 eine vergrößerte Darstellung einer LUDV-Druckwaage gemäß Figur 1;
• Figuren 3 und 4 Ausführungsbeispiele der halbseitig gedämpften Druckwaage aus Figur 1;
• Figur 5 eine LUDV-Druckwaage mit integrierter Lasthaltefunktion;
• Figuren 6 und 7 Betriebszustände der LUDV-Druckwaage aus Figur 5 und
• Figur 8 ein weiteres Ausführungsbeispiel einer LUDV-Druckwaage mit Lasthaltefunktion.

In the following preferred embodiments of the invention will be explained in more detail with reference to schematic drawings. Show it:

• Figure 1 is a sectional view of a valve disc with a half-damped LUDV pressure compensator;
• Figure 2 is an enlarged view of a LUDV pressure compensator according to Figure 1;
• Figures 3 and 4 embodiments of the half-side damped pressure compensator of Figure 1;
• FIG. 5 shows a LUDV pressure balance with integrated load holding function;
• Figures 6 and 7 operating states of the LUDV pressure compensator of Figure 5 and
• Figure 8 shows another embodiment of a LUDV pressure compensator with load-holding function.

Figure 1 shows a section through a valve disc 1 of a control block of a mobile implement, such as a mini or compact excavator, backhoe loader, telescopic loader, skid steer loader. In this Valve disk 1, a proportionally adjustable directional control valve 4 and a LUDV pressure compensator 2 are added, via which the pressure fluid flow between a connected to working ports A, B consumers of the mobile implement and a pressure port and a tank port (both not shown) is controllable. The directional control valve 4 has a pressure medium volume flow to the consumer determining the speed part 6 and two direction parts 8, 10, via which the flow direction of the pressure medium is controlled to or from the consumer.

The directional control valve 4 has a valve slide 12 which is biased by a centering spring 14 in the illustrated basic position. The actuation of the valve spool 12 via a laterally led out of the valve disc 1 operating portion 16 which is articulated to a lever or the like in the driver's cab.

The valve spool 12 is guided in a valve bore 18, which in the radial direction to a pressure chamber 20, an inlet chamber 22, two approximately symmetrically to the pressure chamber 20 disposed drain chambers 24, 25, two on both sides thereof arranged working chambers 26, 28 and two adjacent thereto tank chambers 30, 32 is extended. The valve spool 12 has a central measuring diaphragm collar 34 which, together with the remaining annular web between the pressure chamber 20 and the inlet chamber 22, determines a measuring diaphragm forming the speed part 6. On both sides of this measuring diaphragm collar 34, two control collars 36, 38 and two tank collars 40, 42 of the directional parts 8, 10 are arranged on the valve slide 12.

The pressure chamber 20 is connected to the pressure port P and the two tank chambers 30, 32 are connected to the tank port T. The inlet chamber 22 is connected via a Inlet channel 44 connected to the input of the pressure compensator 2. Their output is connected via two drain channels 46, 48 with the drain chamber 24 and 25, respectively. The two working chambers 26, 28 are connected via working channels 50 and 52 to the working port A and B, respectively.

The structure of the pressure compensator 2 will be explained with reference to the enlarged view in FIG. In Figures 1 and 2, the pressure compensator 2 is shown in the fully open operating position in which the inlet channel 44 is completely opened to the drainage channel 46 out. The pressure compensator 2 has a guided in a pressure compensator bore 54 pressure compensator piston 56, which is designed as a hollow stepped piston and is guided on a correspondingly stepped fixed inner part 58. This is determined in the axial direction by a shoulder 60 of the housing part and a screwed into the pressure balance bore 54 screw plug 62. As can be seen in particular from FIG. 1, to compensate for a constructionally required axial play, the inner part 58 is pretensioned by means of a spring 64 in the direction of the shoulder 60. This spring 64 can not be seen in the partially sectioned illustration in FIG. The inner part 58 further has a blind hole 66 which is closed to the shoulder 60 and which opens into a rear spring chamber 68 which is connected via radial bores 70 with a rear pressure chamber 72, in which the Druckwaagenkolbenendabschnitt larger diameter with its rear annular end face dips. This pressure chamber 72 is acted upon via an LS channel 74 with the greatest load pressure of all connected to the control block consumers.

An inner ring end face 76 defines a damping space 80 with an annular end face 78 of the inner part in the axial direction, which is connected via a peripheral wall of the inner part 58 in the radial direction (perpendicular to the plane) passing through damping nozzle 82 with the blind hole 66. Parallel to this damping nozzle 82, which has a comparatively small diameter, a plurality of radially extending connecting recesses 84 are formed in the inner part 58, which also extend between the blind hole 66 and the damping chamber 80. The damping chamber-side mouth region of the connecting recesses 84 is closed by an elastic O-ring 86, which acts as a check valve, which prevents a pressure medium flow from the damping chamber 80 through the connecting recesses 84 in the blind hole 66 and allows in the opposite direction.

At the foot-side end portion of the inner part 58, an annular groove 88 is formed, in which a Lastmeldeblende 90 opens, via which the entrance of the pressure compensator 2 is connected to the blind hole 66. This load indicator panel 90 is opened when the pressure compensator 2 is fully open, so that the pressure at the inlet of the pressure compensator, i. the individual load pressure also acts in the back pressure chamber 72 and is reported in the LS channel 74. In the closed position of the pressure compensator piston 56, the load indicator shutter 90 is closed in the embodiment shown in Figure 2.

In the illustrated in Figure 1 basic position of the valve spool the metering orifice is closed and shut off the two working ports A, B with respect to the tank channel T. The pressure balance is closed and thus the connection between the channels 46, 48 and 44 shut off. When axial displacement of the valve spool 12, for example, to the right in Figure 1, is formed by the control diaphragm 34 formed on the Meßblendenbund a Orifice opening is opened, via which the pressure chamber 20 is connected to the inlet chamber 22. At the beginning of this Aufsteuerbewegung the pressure in the inlet chamber 44 corresponds approximately to the pump pressure. This pump pressure acts on the outer annular end face 92 of the pressure compensator piston 56 in the opening direction, while the rear-side annular end face 94 is acted upon by the pressure in the pressure chamber 72 and thus the load pressure. The pump control causes the pump pressure to increase until the load pressure, which keeps the pressure compensator closed, is reached. The pressure compensator piston 56 lifts from its stop on the shoulder 60 and opens the connection from the inlet channel 44 to the working channel 46. In this variant shown, the control amount for the connected to the pump control LS-channel is taken by the consumer, which in unfavorable operating conditions of the connected consumer can fall.

In the case in which only one consumer is controlled, the pressure compensator 2 opens completely, so that the load indicator diaphragm 90 is opened and, correspondingly, the load pressure in the working channel 46 is guided into the pressure chamber 72 and thus into the LS channel 74.

During opening movements of the pressure compensator piston 56, pressure medium must be displaced from the decreasing damping chamber 80. Since the comparatively large cross-section of the connecting recesses 84 are shut off by the O-ring 86, the pressure medium flows through the small damping nozzle 82 into the blind hole bore 66, so that the opening movement of the pressure compensator piston 56 is relatively strongly damped.

If a second consumer is actuated with a higher load pressure, then this higher load pressure acts in all Consumers common LS channel 74 - the pressure compensator piston 56 is moved accordingly in the closing direction until a pressure equilibrium is established. In this control position, the pressure drop across the associated metering orifice is kept constant, whereby the flow rate selected for each consumer is kept relatively constant.

During this closing movement of the pressure compensator piston 56, the damping chamber 80 is increased, so that accordingly pressure medium can flow from the blind bore 66 into the damping chamber 80. The elasticity of the O-ring 86 allows a pressure medium flow in this direction, so that pressure medium can flow through the comparatively large cross-section of the connecting recesses 84 - the closing movement of the damping piston is almost undamped, so that the load-higher consumer is driven virtually without delay.

Two variants of a pressure compensator 2 are shown in FIGS. 3 and 4, wherein instead of the O-ring 86, other non-return valve arrangements are used.

The basic structure of the pressure compensator 2 is in each case the same as in FIG. 2, so that only the differences are discussed below. In the embodiment shown in Figure 3, the connecting recesses 84 are not formed in the radial direction between the damping chamber 80 and the blind hole 66, but they are designed as a trained symmetrically to the pressure compensator axis bore star. About these axially extending connecting recesses 84 of the rear pressure chamber 72 is connected directly to the damping chamber 80. The check valve is formed by an annular closing disc 96, the the inner part 58 engages around and is inserted into an axial groove 98 on the lower end face of the larger end section of the inner part 58 in FIG. The closing disc 96 is biased in the closing direction by the force of a valve spring 100, which is supported on a spring plate 102 inserted into an annular groove of the inner part 58. The strength of the valve spring 100 is selected so that a pressure medium flow from the rear pressure chamber 72 into the damping chamber 80 during the closing movement of the pressure compensator piston 56 can be carried out with a relatively low pressure loss, so that the damping is substantially lower than during the closing movement of the pressure compensator piston, in the the pressure medium has to flow over the small damping nozzle 82.

In the embodiment shown in FIG. 4, instead of the bore star which can be closed by the valve disc 96, a single axial bore is provided in the inner part, into which a check valve 104 having a valve body 106 which is biased against a valve seat 108 is inserted. The function of this check valve 104 corresponds to that of the above-described embodiment, so that further explanations are unnecessary.

FIG. 5 shows a further variant of a LUDV pressure compensator 2 according to the invention, in which, in addition to the one-sided damping described above, a load-holding function is integrated which prevents the load from sagging, so that additional load-holding valves can be dispensed with.

The basic structure of the embodiment shown in Figure 5 largely corresponds to that the above-described embodiments, so that only the differences will be discussed.

Also in the variant according to FIG. 5, a pressure balance piston 56 is guided axially displaceably on an inner part 58. The rear annular end face 94 is the pressure in the pressure chamber 72 and the outer annular end face 92 of the pressure at the entrance of the pressure compensator 2, i. from the pressure in the inlet channel 44 (downstream of the metering orifice) acted upon. In the interior of the pressure compensator piston 56, in turn, the damping chamber 80 is formed, so that the inner ring end face 76 is acted upon by the pressure in this damping chamber 80 in the closing direction. Between the blind bore 66 of the inner part 58 and the damping chamber 80 are - as in the embodiment of Figure 2 - radially extending connecting recesses 84 are formed, the damping chamber side of an O-ring 86 are closed. At the foot end portion, the inner part 58 has a Lastmeldeblende 90. Up to here, the embodiment of FIG 5 completely corresponds to the embodiment of Figure 2. The main difference is that the small damping nozzle 82 is not formed in the inner part but in the mantle of the damping piston 56, so that the damping chamber 80 is not connected via this damping nozzle 82 with the blind hole 66 but with the working channels 46, 48. That By means of the damping nozzle 82, the load pressure acting on the associated load acts in the damping chamber 80.

Furthermore, a bore 110 is formed at the end portion of the pressure compensator piston 56 with a smaller diameter, which in the closed position of the pressure compensator 2 shown in FIG Lastmeldeblende 90 is aligned so that pressure medium from the inlet channel 44 can enter the blind hole 66.

The damping chamber 80 further acts as a spring chamber for a spring 112, which is supported on the adjacent annular end face of the inner part 58 and engages the inner ring end face 76 of the pressure compensator piston 56. Also, this spring 112 serves to compensate for the design predetermined clearance in the axial direction and to ensure a rapid closing of the pressure compensator piston 56 - in principle, could be dispensed with the spring 112.

In the basic position of the spool valve and closed pressure compensator 2, the load pressure on the associated load via the working channels 46, 48 and the small damping nozzle 82 in the damping chamber 80 acts. The O-ring 86 blocks the passage to the blind hole 66 from. In the blind bore 66 and in the associated pressure chamber, the pressure in the inlet channel 44 is effective via the load indicator cover 90 and the bore 110.

Upon actuation of the valve spool 12, this pressure corresponds to the inlet channel 44, i. the pressure downstream of the metering orifice initially substantially the pump pressure, so that in the pressure chamber 72 also applies pump pressure. In this embodiment, the LS-channel 74 is thus filled via the pump in the illustrated basic position of the pressure balance and not - as in the above embodiments - on the load, so that a sagging of the consumer in the control due to a filling of the LS channel 74th is prevented.

The pump control of the pump, not shown, causes the applied pump pressure to rise until the load pressure which keeps the pressure compensator closed is reached. Since at the beginning of the control in the LS channel 74, the pump pressure acts and this is reported to the pump controller, this retracts, so to speak, "even high" until the balance of power is achieved with the effective force in the closing direction, which essentially by the on the inner ring end face 76 acting load pressure (and the pressure in the rear pressure chamber) is determined. The pressure compensator piston 56 then begins to open the passage to the working channel 46, 48 and thus to the consumer. At the same time, the overlap of the load indicator panel 90 with the bore 110 is released, so that the load indicator panel 90 is closed.

This operating state is shown in FIG. At first, a very low pressure medium volume flow flows to the consumer, ie the pressure drop across the metering orifice is low. The regulated by the pump control pressure drop is almost completely over the pressure compensator, which is further opened due to this pressure difference. The pressure compensator is finally completely opened (see FIG. 7), wherein the load alarm shutter 90 is opened again by the lower annular surface 114 of the pressure compensator piston 56. About the load indicator panel 90, the blind hole 66, the pressure chamber 72 and thus the LS channel 74 is supplied with a volume flow which is substantially constant by a downstream current regulator. The pressure drop, which generates this volume flow between the front and back of the pressure compensator 2, is greater than the force of the spring 112 - the pressure compensator remains fully open. The spring is - as I said - only to keep the pressure compensator in readiness to close.

Now, if another consumer is operated with a higher load pressure, the pressure balance of the first driven consumer is brought in the manner described above in its control position, so that the pressure drop across the metering orifice remains constant and all consumers are supplied load-independent with pressure medium.

If the pump pressure drops due to fluctuations in the pressure medium supply under the load pressure, the pressure compensator piston 56 is moved quickly by the force acting on its inner ring end face 76 load pressure in its closed position and acts as a load-holding valve.

FIG. 8 finally shows a variant of the exemplary embodiment described in FIGS. 5 to 7, in which no radial bore 110 is formed at the smaller diameter of the hollow pressure balance piston 56 but recesses 116 are formed in the end surface formed by the annular surface 114, which open into an annular gap 118. which is formed by a downgrading of the inner part 58. This annular gap 118 extends in the axial direction to the load indicator panel 90. When completely open pressure compensator (left in Figure 8), the load alarm panel 90 is fully opened, so that no hydraulic resistance (annular gap 118) is connected upstream.

In this variant, thus the load-sensing line of the control block is supplied via all disks with pressure medium, which is tapped by the pump. Preliminary tests showed that this variant influences the LUDV control characteristics, since the LS line is supplied by all active consumers.

The applicant reserves the right to focus on the load holding function of a separate patent application, the claim may be directed to the application of the damping chamber 80 with the load pressure.

Disclosed is a hydraulic control arrangement for load-independent control of a consumer with a continuously adjustable directional control valve and a downstream pressure compensator. According to the invention, the pressure compensator is damped on one side, so that the movement is damped in the opening direction and the movement in the closing direction is substantially unattenuated. It is further disclosed a control arrangement in which in the pressure compensator a load-holding function is integrated.

LIST OF REFERENCE NUMBERS

1

valve disc

2

LUDV pressure

4

way valve

6

speed part

8th

direction part

10

direction part

12

valve slide

14

centering

16

actuating section

18

valve bore

20

pressure chamber

22

inlet chamber

24

drain chamber

25

drain chamber

26

working chamber

28

working chamber

30

tank chamber

32

tank chamber

34

Measuring orifice collar

36

control collar

38

control collar

40

tank Bund

42

tank Bund

44

inlet channel

46

drain channel

48

drain channel

50

working channel

52

working channel

54

Pressure compensator bore

56

Pressure regulator piston

58

inner part

60

shoulder

62

Screw

64

feather

66

Blind hole

68

spring chamber

70

radial bore

72

pressure chamber

74

LS-channel

76

Inner annular face

78

Annular face

80

damping space

82

damping

84

connecting recess

86

O-ring

88

ring groove

90

Load-detecting orifice

92

outer ring end face

94

rear ring end face

96

closing disk

98

axial groove

100

valve spring

102

spring plate

104

check valve

106

valve body

108

valve seat

110

drilling

112

feather

114

ring surface

116

recesses

118

annular gap

Claims (15)

1. A hydraulic control arrangement for the load-independent control of a consumer, comprising a continuously adjustable distribution valve (4) which forms a metering orifice having a pressure compensator (2) down the line the pressure compensator piston (56) of which is a step piston so that the pressure compensator (2) includes a rear pressure chamber (72) and an annular damping chamber (80) which is connected via a damping nozzle (82) to an adjacent chamber (66, 46; 48) guiding pressure medium, wherein a control pressure acting on the pressure compensator piston (56) in the closing direction can be applied to the pressure chamber (72) and to the damping chamber (80), and wherein a pressure downstream of the metering orifice in the opening direction is applied to an outer annular face (92) of the pressure compensator piston (56), characterized by a connecting recess (84) between the rear pressure chamber (72) and the damping chamber (80), to which a check valve (86, 986, 104) opening toward the damping chamber (80) is allocated.
2. A control arrangement according to claim 1, wherein the pressure compensator piston (56) is a hollow piston and is guided on an axial male member (58) having a blind hole bore (66) which opens into the rear pressure chamber (72).
3. A control arrangement according to claim 2, wherein the damping nozzle (82) connects the damping chamber (80) to a passage (46, 48) guiding the load pressure of the corresponding consumer.
4. A control arrangement according to claim 2, wherein the damping nozzle (82) connects the damping chamber (80) to the rear pressure chamber (72).
5. A control arrangement according to any one of the claims 2 to 4, wherein at a bottom-side end portion of the male member (58) a load-detecting orifice (90) opening into the blind hole bore (66) is provided which is controlled to be completely opened in the opening position of the pressure compensator piston (56).
6. A control arrangement according to claim 5, wherein at the smaller diameter of the pressure compensator piston (56) a bore (110) or a circumferential recess (116) is formed via which the pressure downstream of the metering orifice can be signaled into the blind hole bore (66).
7. A control arrangement according to claim 6, wherein in the closing position of the pressure compensator piston (56) the bore (110) is in overlapping with the load-detecting orifice (90) which can be controlled to be closed in the subsequent stroke of the pressure compensator piston (56) and can be controlled to be opened again by the pressure compensator piston (56) upon reaching the opening position.
8. A control arrangement according to claim 7, wherein the circumferential recess (116) opens into an annular gap (118) between the male member (58) and the pressure compensator piston (56) extending toward the load-detecting orifice (90).
9. A control arrangement according to any one of the preceding claims, wherein the connecting recess (84) is formed by a bore star opening into the blind hole bore (66) and being closable by an O-ring (86) placed on the male member (58).
10. A control arrangement according to any one of the claims 1 to 8, wherein the connecting recess (84) is formed by a bore of the male member (58) opening into the pressure chamber (80) in which bore a check valve (104) is accommodated.
11. A control arrangement according to any one of the claims 1 to 9, wherein the pressure compensator piston (56) is biased into its closing position by a spring.
12. A pressure compensator for a hydraulic control arrangement according to any one of the preceding claims, comprising a stepped pressure compensator piston (56) in the form of a hollow piston and guided on a male member (58), whose rear annular face (94) delimits a rear pressure chamber (72) and whose inner annular face (76) delimits an annular damping chamber (80) in sections, said damping chamber being connected to an adjacent chamber guiding pressure medium via a damping nozzle (82), wherein a pressure acting in the closing direction can be applied to the inner annular face (76) and the rear annular face (94) and a pressure acting in the opening direction can be applied to an outer annular face (92), characterized by a connecting recess (84) of the male member (58) between the rear pressure chamber (72) and the damping chamber (80), a check valve (86, 96, 104) opening toward the damping chamber (80) being allocated to said connecting recess.
13. A pressure compensator according to claim 12, wherein the damping nozzle (82) connects the damping chamber (80) to a pressure chamber guiding the load pressure of a corresponding consumer.
14. A pressure compensator according to claim 12, wherein the damping nozzle (82) connects the damping chamber (80) to the rear pressure chamber (72).
15. A pressure compensator according to any one of the claims 12 to 14, wherein in the male member a blind hole bore (66) opening into the rear pressure chamber (72) is formed into which a load-detecting orifice (90) opens at the bottom side, a bore (110) of the pressure compensator piston (56) being allocated to the load-detecting orifice (90) and being overlapped with the load-detecting orifice (90) in the closing position of the pressure compensator piston (56), wherein upon a subsequent opening movement the load-detecting orifice (90) can be controlled to be closed and in the completely opened position of the pressure compensator piston (56) can be controlled to be opened again.

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