EP1629209A1

EP1629209A1 - Hydraulic control arrangement

Info

Publication number: EP1629209A1
Application number: EP04735220A
Authority: EP
Inventors: Wolfgang Kauss; Didier Desseux
Original assignee: Bosch Rexroth AG
Current assignee: Bosch Rexroth AG
Priority date: 2003-06-04
Filing date: 2004-05-28
Publication date: 2006-03-01
Anticipated expiration: 2024-05-28
Also published as: JP4851318B2; DE502004002559D1; DE10325296A1; US7628174B2; ATE350586T1; JP2006526746A; EP1629209B1; US20060191582A1; WO2004109125A1

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

description

Hydraulic control arrangement

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 AI. In such a LUDV system, an adjustable metering orifice with a downstream pressure compensator is assigned to each consumer, the latter keeping the pressure drop across the metering orifice constant, so that the amount of pressure medium 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, a large number of such valve arrangements are connected in parallel in the case of mobile work tools, the system's individual pressure compensators ensure that in the event that a hydraulic pump of the system has been adjusted to the maximum stroke volume and the pressure medium flow is insufficient, the predetermined pressure drop exceeds to maintain the metering orifices of the valve arrangements assigned to each consumer, 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. Due to this flow pressure-independent flow distribution (LUDV), all actuated consumers move at a speed that is reduced by the same percentage. LUDV hydraulic systems of this type are used to an ever increasing extent in mobile work devices with combined movements. The working movements of these mobile working devices (mini and compact excavators, backhoe loaders, telescopic loaders, compact loaders, etc.) should follow the control by the driver without vibration and pressure. It has been shown that damping the LUDV pressure compensators is necessary for vibration-free control.

Damping is known for example from US 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, while the pressure applied downstream of the measuring orifice, as a rule the load pressure of the controlled one, is applied to a front end face of the pressure compensator piston Consumer, acts in the opening direction. A damping nozzle is provided between the rear pressure chamber and the damping chamber, through which the pressure medium must flow out of or into the damping chamber during the axial displacement of the pressure compensator piston, so that the pressure compensator piston movement is damped. Such damping inevitably leads to delays in opening and closing the pressure compensator with the consequence of a delayed start of movement of work movements with high load.

In contrast, the invention has for its object to provide a control arrangement and a suitable LUDV pressure compensator, in which the delay in the work movement of a consumer is minimized despite damping the pressure compensator. This object is achieved with regard to the hydraulic control arrangement by the features of claim 1 and with regard to the pressure compensator by the features of the independent claim 12.

According to the invention, in addition to the damping nozzle connecting the damping chamber to 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 check valve opening in the direction of the damping chamber. As a result of this measure, the movement of the pressure compensator piston in the opening direction as a function of the orifice cross section is damped relatively strongly, while in the closing direction the check valve opens and thus opens a comparatively large cross section - i.e. the pressure compensator is damped on one side, so that the pressure compensator of a consumer with low load pressure closes quickly, for example, and in this way the rapid pressure build-up to a higher one

Allows load pressure in another disc.

In a preferred embodiment, the pressure compensator piston is designed as a stepped hollow piston, as described in US Pat. No. 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. An inner ring end face formed by the stepped part of the pressure compensator piston delimits the damping space with a correspondingly designed section of the inner part. The foot-side ring end face of the step piston is acted upon by the pressure downstream of the measuring orifice in the opening direction of the pressure compensator. In the known solutions, a rear control chamber of the pressure compensators is connected to the load signaling line, in which the highest load pressure of all controlled consumers, which is tapped via a shuttle valve chain, is present. If the load pressure of an actuated hydraulic consumer quickly rises above the currently prevailing highest load pressure, the pressure at the front of the pressure compensator piston of the associated pressure compensator increases immediately, while a corresponding pressure increase in the rear control chamber takes place with a delay via the shuttle valve chain and the load signaling line. The resulting short-term force imbalance on the control piston of the pressure compensator can have a negative effect on the control of the hydraulic consumer. For example, the hydraulic consumer can sag briefly or the load-independent flow distribution can be disrupted.

To avoid such an undesirable 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 from the consumer can be prevented via the pressure compensator. However, such additional load holding valves make the control arrangement more expensive and require a considerable amount of space.

To overcome this disadvantage, pressure balances 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 holding is integrated in the function of the pressure compensator. The pressure compensator is provided with two pressure compensator pistons arranged one behind the other, which are switched in such a way that the pressure compensator is closed when the pressure at the inlet of the pressure compensator is open Pressure compensator piston is lower than the individual load pressure.

DE 40 05 966 C2 proposes a solution in which a shuttle valve is integrated in the pressure compensator piston, by means of which the pressure downstream of the metering orifice and in the load signaling channel is compared and reported to the rear control chamber.

DE 296 17 735 Ul describes a pressure compensator in which the load message is complex

Shuttle valve switching with check valves and nozzles takes place in order to keep the pressure balance of the load holding function in the closed state.

All the known solutions described above with a load holding function in the pressure compensator have the disadvantage in common that a considerable effort is required to tap a control pressure which

Pressure compensator piston in the load holding function in the closing direction.

According to one 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 event that the pressure at the inlet of the pressure compensator drops below this load pressure, the pressure compensator piston moves via the pressure sensor Damping space adjacent individual load pressure is brought into its closed position, so that the pressure compensator takes over the load holding function. Compared to 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 can also connect the damping chamber to the rear pressure chamber, although the load holding function is dispensed with.

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

In a particularly preferred exemplary embodiment, a bore or a recess is formed on the smaller diameter of the pressure compensator piston, which can be positioned with respect to the transverse bore in such a way that the pressure downstream of the measuring orifice is reported into 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 connection is opened only during the initial stroke (seen from the closed position) of the pressure compensator and when the pressure compensator is fully open, while in the area in between this connection is closed, so that the rear pressure chamber is then subjected to the highest effective load pressure , while at the start of opening the pressure compensator with the pressure downstream of the

Orifice plate - i.e. about the pump pressure - is applied.

The check valve according to the invention can be formed by a simple O-ring on the

Inner part is put on or by one into one Prestressed locking disc in the closed position. Alternatively, conventional check valves with spring-loaded closing bodies can also be used.

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

Other advantageous developments of the invention are the subject of further dependent claims.

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

Figure 1 is a sectional view of a valve disc with a half-sided damped LUDV pressure compensator;

FIG. 2 shows an enlarged illustration of an LUDV pressure compensator according to FIG. 1;

Figures 3 and 4 embodiments of the half-sided damped pressure compensator from Figure 1;

FIG. 5 shows an LUDV pressure compensator with an integrated load holding function;

Figures 6 and 7 operating states of the LUDV pressure compensator from Figure 5 and

Figure 8 shows another embodiment of an LUDV pressure compensator with load holding function.

Figure 1 shows a section through a valve disc

1 of a control block of a mobile working device, for example a mini or compact excavator,

Backhoe loaders, telescopic handlers, skid steer loaders. In this Valve disc 1 includes a proportionally adjustable directional valve 4 and an LUDV pressure compensator 2, via which the pressure medium flow between a consumer of the mobile working device connected to work connections A, B and a pressure connection and a tank connection (both not shown) can be controlled. The directional control valve 4 has a speed part 6 determining the pressure medium volume flow to the consumer and two directional parts 8, 10, via which the direction of flow of the pressure medium to and from the consumer is controlled.

The directional control valve 4 has a valve spool 12, which is shown in FIG

Basic position is biased. The valve spool 12 is actuated laterally from the

Valve disc 1 led out operating section 16 which is hinged to an actuating lever or the like in the driver's cabin.

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

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 via a Inlet channel 44 connected to the inlet of the pressure compensator 2. Its outlet is connected to the outlet chamber 24 and 25 via two outlet channels 46, 48. The two working chambers 26, 28 are connected to the working connection A and B via working channels 50 and 52, respectively.

The structure of the pressure compensator 2 is explained on the basis of the enlarged illustration in FIG. 2. 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 outlet channel 46. The pressure compensator 2 has a pressure compensator piston 56 which is guided in a pressure compensator bore 54 and is designed as a hollow stepped piston and is guided on a correspondingly graduated fixed inner part 58. This is fixed in the axial direction by a shoulder 60 of the housing part and a screw plug 62 screwed into the pressure compensator bore 54. As can be seen, in particular, in FIG

The inner part 58 is biased axially in the direction of the shoulder 60 by means of a spring 64. This spring 64 cannot be seen in the partially sectioned illustration in FIG. 2. The inner part 58 also has a blind hole 66, which is closed towards the shoulder 60 and which opens into a rear spring chamber 68, which is connected via radial bores 70 to a rear pressure chamber 72, into which the pressure compensator piston end section with the larger diameter with its rear ring end face dips. This pressure chamber 72 is acted upon by an LS channel 74 with the greatest load pressure of all consumers connected to the control block.

An inner ring end face 76 delimits a damping space 80 with an annular end face 78 of the inner part in the axial direction, which is connected to the blind bore 66 via a damping nozzle 82 passing through a circumferential wall of the inner part 58 in the radial direction (perpendicular to the plane of the drawing). Parallel to this damping nozzle 82, which has a comparatively small diameter, a plurality of radially extending connecting recesses 84 are also formed in the inner part 58, which also extend between the blind hole 66 and the damping space 80. The opening region of the connection recesses 84 on the damping chamber side 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 connection recesses 84 into the blind hole 66 and allows it in the opposite direction.

On the foot-side end section of the inner part 58, an annular groove 88 is formed, into which a load-reporting diaphragm 90 opens, via which the inlet of the pressure compensator 2 is connected to the blind hole 66. This load-sensing orifice 90 is opened when the pressure compensator 2 is fully open, so that the pressure at the inlet of the pressure compensator, i.e. the individual load pressure also acts in the rear pressure chamber 72 and is reported in the LS channel 74. In the closed position of the pressure compensator piston 56, the load-sensing diaphragm 90 is closed in the exemplary embodiment shown in FIG. 2.

In the basic position of the valve slide shown in FIG. 1, the measuring orifice is closed and the two working connections A, B are blocked off from the tank channel T. The pressure compensator is closed and thus the connection between the channels 46, 48 and 44 is blocked. When the valve slide 12 is axially displaced, for example to the right in FIG. 1, the control notches formed on the orifice collar 34 result in a notch Measuring aperture opening opened, via which the pressure chamber 20 is connected to the inlet chamber 22. At the beginning of this opening movement, the pressure in the inlet chamber 44 corresponds approximately to the pump pressure. This pump pressure acts on the outer ring end face 92 of the pressure compensator piston 56 in the opening direction, while the rear ring 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 rise 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 quantity for the LS channel connected to the pump control is taken from the consumer, which means that the connected consumer in unfavorable operating conditions can sink.

In the case in which only one consumer is actuated, the pressure compensator 2 opens completely, so that the load-sensing diaphragm 90 is opened and the load pressure in the working channel 46 is accordingly led into the pressure chamber 72 and thus into the LS channel 74.

During opening movements of the pressure compensator piston 56, pressure medium has to be displaced from the shrinking damping space 80. Since the comparatively large cross-section of the connecting recesses 84 is blocked by the O-ring 86, the pressure medium flows through the small damping nozzle 82 into the blind hole 66, so that the opening movement of the pressure compensator piston 56 is damped relatively strongly.

If a second consumer with a higher load pressure is actuated, 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 assigned orifice plate is kept constant, which means that the flow rate selected for each consumer is kept constant at the same ratio.

During this closing movement of the pressure compensator piston 56, the damping chamber 80 is enlarged, so that pressure medium can correspondingly flow from the blind hole 66 into the damping chamber 80. The elasticity of the O-ring 86 leaves a pressure medium flow in it

Direction to, so that pressure medium due to the comparatively large cross section of the

Connection recesses 84 can flow - the

The closing movement of the damping piston is almost undamped, so that the higher-load consumer is controlled with practically no delay.

FIGS. 3 and 4 show two variants of a pressure compensator 2, other check valve arrangements being used instead of the O-ring 86.

The basic structure of the pressure compensator 2 is in each case the same as in FIG. 2, so that only the differences are dealt with below. In the exemplary embodiment shown in FIG. 3, the connecting recesses 84 are not designed in the radial direction between the damping space 80 and the blind hole 66, but they are designed as a bore star that is symmetrical to the pressure compensator axis. The rear pressure chamber 72 is connected directly to the damping chamber 80 via these axially extending connecting recesses 84. The check valve is formed by an annular closing disk 96 which encompasses the inner part 58 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. 3. The closing disk 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 in the inner part 58. The strength of the valve spring 100 is selected such that a pressure medium flow from the rear pressure chamber 72 into the damping chamber 80 can take place with a comparatively small pressure loss during the closing movement of the pressure compensator piston 56, so that the damping is substantially less than during the closing movement of the pressure compensator piston, in which the pressure medium must flow through the small damping nozzle 82.

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

FIG. 5 shows a further variant of an LUDV pressure compensator 2 according to the invention, in which, in addition to the above-described one-sided damping, a load holding function is also integrated which prevents the

Load prevents, so that additional load holding valves can be dispensed with.

The basic structure of the embodiment shown in Figure 5 largely corresponds to that of the above-described exemplary embodiments, so that only the differences are dealt with.

In the variant according to FIG. 5, too, a pressure compensator piston 56 is axially displaceably guided on an inner part 58. The rear ring face 94 is from the pressure in the pressure chamber 72 and the outer ring face 92 from the pressure at the inlet of the pressure compensator 2, i.e. acted upon by the pressure in the inlet channel 44 (downstream of the orifice plate). The damping chamber 80 is in turn formed in the interior of the pressure compensator piston 56, so that the pressure in this damping chamber 80 acts on the inner ring end face 76 in the closing direction. Between the blind hole 66 of the inner part 58 and the damping space 80, as in the exemplary embodiment according to FIG. 2, radially extending connecting recesses 84 are formed, which are closed by an O-ring 86 on the damping space side. At the foot-side end section, the inner part 58 has a load-sensing diaphragm 90. Up to this point, the exemplary embodiment according to FIG. 5 corresponds completely to the exemplary embodiment according to FIG. 2. The main difference is that the small damping nozzle 82 is not formed in the inner part but in the jacket of the damping piston 56, so that the damping chamber 80 is connected via this damping nozzle 82 not to the blind hole 66 but to the working channels 46, 48. That The load pressure effective at the associated consumer acts in the damping chamber 80 via the damping nozzle 82.

Furthermore, at the end portion of the

Pressure compensator piston 56 with a smaller diameter

Bore 110 formed, which in the closed position shown in Figure 5 of the pressure compensator 2 with the Load-sensing diaphragm 90 is aligned so that pressure medium can enter the blind bore 66 from the inlet channel 44.

The damping chamber 80 also acts as a spring chamber for a spring 112 which is supported on the adjacent ring end face of the inner part 58 and which engages on the inner ring end face 76 of the pressure compensator piston 56. This spring 112 also serves to compensate for the play in the axial direction and to ensure that the pressure compensator piston 56 closes quickly - in principle, the spring 112 could be dispensed with.

In the basic position of the control slide and with the pressure compensator 2 closed, the load pressure acts on the associated consumer via the working channels 46, 48 and the small damping nozzle 82 in the damping chamber 80. The O-ring 86 blocks the passage to the blind hole 66. The pressure in the inlet channel 44 is effective in the blind hole 66 and in the pressure chamber connected to it via the load-sensing diaphragm 90 and the hole 110.

When the valve slide 12 is actuated, this pressure corresponds to the inlet channel 44, i.e. the pressure downstream of the metering orifice is initially essentially the pump pressure, so that pump pressure is also present in the pressure chamber 72. In this exemplary embodiment, in the illustrated basic position of the pressure compensator, the LS channel 74 is filled via the pump and not - as in the previously described exemplary embodiments - via the load, so that the consumer sags during activation due to the LS channel 74 being filled is prevented.

The pump control of the pump, not shown, allows the pump pressure to rise until the load pressure, which keeps the pressure compensator closed, has been reached. Since the pump pressure acts in the LS channel 74 at the start of the control and this is reported to the pump controller, it so to speak "pulls itself up" until the force equilibrium is reached with the force acting in the closing direction, which is essentially due to the area on the inner ring 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 alarm panel 90 with the bore 110 is removed, so that the load alarm panel 90 is closed.

This operating state is shown in FIG. 6. A very low pressure medium volume flow initially flows to the consumer, i.e. the pressure drop across the orifice plate is low. The pressure drop regulated by the pump control still arises almost completely above the pressure compensator, which is due to this

Pressure difference is opened further. Finally, the pressure compensator is opened completely (see FIG. 7), the load-sensing diaphragm 90 being opened again through the lower annular surface 114 of the pressure compensator piston 56. Via the load-sensing diaphragm 90, the blind hole 66, the pressure chamber 72 and thus the LS channel 74 are now supplied with a volume flow which is essentially constant by a downstream flow regulator. The pressure drop that this volume flow generates between the front and rear of the pressure compensator 2 is greater than the force of the spring 112 - the pressure compensator remains completely open. As I said, the spring only serves to keep the pressure compensator in the ready position to close. If another consumer with a higher load pressure is now actuated, the pressure compensator of the consumer controlled first is brought into its control position in the manner described above, so that the pressure drop across the measuring orifice remains constant and all consumers are supplied with pressure medium regardless of the load.

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

Finally, FIG. 8 shows a variant of the exemplary embodiment described in FIGS. 5 to 7, in which, on the smaller diameter of the hollow pressure compensator piston 56, there are no radial bores 110, but recesses 116 are formed in the end face formed by the annular surface 114, which recesses are formed in an annular gap

118 open, which is formed by a downgrading of the inner part 58. This annular gap 118 extends in the axial direction up to the load-sensing orifice 90. When the pressure compensator is fully open (left in FIG. 8), the load-sensing orifice 90 is fully opened, so that no hydraulic resistance (annular gap 118) is connected upstream.

In this variant, the load reporting line of the control block is thus supplied with pressure medium via all disks, 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 direct its own patent application to the load holding function, whereby the claim can be directed to the damping space 80 being subjected to the load pressure.

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

Reference symbol list:

Valve disc LUDV pressure compensator Directional control valve Speed section Directional part Directional valve Valve slide Centering spring arrangement Actuating section Valve bore Pressure chamber Inlet chamber Drain chamber Drain chamber Work chamber Work chamber Tank chamber Tank chamber Measuring orifice collar Control collar Control collar Tank collar Tank collar Feed channel Drain channel Drain channel Working channel Working channel Pressure compensator bore Pressure compensator piston Inner part Shoulder Locking screw feather

Blind hole

spring chamber

radial bore

pressure chamber

LS-channel

Inner ring end face

Annular face

damping space

damping

connecting recess

O-ring

ring groove

Load-sensing panel outer ring face rear ring face

closing disk

axial groove

valve spring

spring plate

check valve

valve body

valve seat

drilling

feather

ring surface

recesses

annular gap

Claims

claims

1. Hydraulic control arrangement for load-independent control of a consumer, with a continuously adjustable directional valve (4), which forms a measuring orifice, which is followed by a pressure compensator (2), the pressure compensator piston (56) of which is designed as a step piston, so that the pressure compensator (2) a rear pressure chamber (72) and an annular one

Damping chamber (80) has a damping nozzle

(82) with an adjacent, pressure medium leading

Space (66, 46; 48) is connected, the pressure space

(72) and the damping chamber (80) can be acted upon by a control pressure acting on the pressure compensator piston (56) in the closing direction, and wherein an outer annular end face (92) of the pressure compensator piston (56) is pressurized downstream of the measuring orifice in the opening direction, characterized by a connecting recess (84) between the rear pressure chamber (72) and the damping chamber (80), which is assigned a check valve (86, 986, 104) opening towards the damping chamber (80).

2. Control arrangement according to claim 1, wherein the pressure compensator piston (56) is a hollow piston and is guided on an axial inner part (58) which has a blind hole {66) which opens into the rear pressure chamber (72).

3. Control arrangement according to claim 2, wherein the damping nozzle (82) connects the damping chamber (80) with a channel carrying the load pressure of the associated consumer (46, 48).

4. Control arrangement according to claim 2, wherein the damping nozzle (82) connects the damping chamber (80) with the rear pressure chamber (72).

5. Control arrangement according to one of the claims 2 to 4, wherein at a foot-side end portion of the inner part (58) in the blind hole (66) opening load indicator orifice (90) is provided, which is fully opened in the opening pitch of the pressure compensating piston (56).

6. Control arrangement according to claim 5, wherein a bore (110) or a circumferential recess (116) is formed on the smaller diameter of the pressure compensator piston (56), via which the pressure downstream of the measuring orifice in the blind hole (66) can be reported.

7. Control arrangement according to claim 6, wherein in the closed position of the pressure compensating piston (56), the bore (110) overlaps with the load-sensing diaphragm

(90) stands, which can be actuated during the subsequent stroke of the pressure compensator piston (56) and can be opened again by the pressure compensator piston (56) when the opening position is reached.

8. Control arrangement according to claim 7, wherein the circumferential recess (116) in an annular gap extending to the load-sensing diaphragm (90)

(118) opens between the inner part (58) and the pressure compensator piston (56).

9. Control arrangement according to one of the preceding claims, wherein the connecting recess (84) is formed by a bore star, which opens into the blind hole (66) and by one the inner part (58) attached O-ring (86) can be closed.

10. Control arrangement according to one of claims 1 to 8, wherein the connecting recess (84) is formed by a bore in the pressure chamber (80) of the inner part (58), in which a check valve (104) is received.

11. Control arrangement according to one of claims 1 to 9, wherein the pressure compensator piston (56) is biased by a spring (112) in its closed position.

12. Pressure compensator for a hydraulic control arrangement according to one of the preceding claims, with a graduated on a inner part (58), designed as a hollow piston pressure compensator piston (56), the rear annular end face (94) of a rear pressure chamber (72) and the inner annular end face (76 ) partially delimits an annular damping space (80) which is connected via a damping nozzle (82) to an adjacent pressure medium-carrying space, the inner ring end face (76) and the rear ring end face (94) having a pressure effective in the closing direction and an outer ring end face (92) can be acted upon by a pressure effective in the opening direction, characterized by a connecting recess (84) in the inner part (58) between the rear pressure chamber (72) and the damping chamber (80), which has a check valve that opens in the direction of the damping chamber (80) (86, 96, 104).

13. Pressure compensator according to claim 12, wherein the damping nozzle (82) the damping chamber (80) with a connects the load pressure of an associated consumer leading pressure chamber.

14. Pressure compensator according to claim 12, wherein the damping nozzle (82) connects the damping chamber (80) with the rear pressure chamber (72).

15. Pressure compensator according to one of the claims 12 to 14, wherein in the inner part a blind hole (66) opening in the rear pressure chamber (72) is formed, in which a load-sensing orifice (90) opens on the foot side, which bore (110) of the pressure compensator piston (56) is assigned, which is in the closed position of the pressure compensating piston (56) in overlap with the load-sensing diaphragm (90), the load-sensing diaphragm (90) being controllable during a subsequent opening movement and being able to be opened again in the fully open position of the pressure compensating piston (56).