EP1781952B1 - Hydraulic control arrangement - Google Patents

Abstract

The invention relates to a hydraulic control arrangement and a pilot-operated pressure relief valve therefor. Said hydraulic control arrangement comprises a differential cylinder provided with a pressure chamber on the piston rod side thereof and another pressure chamber at the bottom thereof which can be connected to a pump or a tank by means of a control valve arrangement in order to actuate the differential cylinder. The pressure in a pressure chamber is defined by a pilot-operated pressure relief valve provided with a pressure switching stage by which means the pressure regulated by the pressure relief valve can be lowered according to the pressure in the other pressure chamber.

Description

The invention relates to a hydraulic control arrangement with a differential cylinder according to the preamble of patent claim 1 and suitable for such a control arrangement pilot-operated pressure relief valve.

Such control arrangements are used in particular in mobile work equipment, for example, to pivot a blade of a wheel loader. In this case, by extending a piston rod of a differential cylinder of the control arrangement, the blade is pivoted down to empty, for example, material received therein. To pick up the material, the piston rod of the differential cylinder is retracted so that the blade is pivoted upwards, ie, away from the ground. Such a solution is for example in the US 4,194,436 described. The control of the differential cylinder via a control valve, which is followed by a boost valve. For retracting the differential cylinder (swinging back of the blade), the control valve and the boost valve are brought into a position in which a pump of the control arrangement with a piston rod-side annulus and a bottom-side cylinder chamber are connected to a tank. To extend the control valve and the boost valve are adjusted so that the cylinder chamber is connected to the pump and the piston rod side annulus is also in communication with the cylinder chamber, so that the displaced from this pressure medium is additionally guided into the cylinder chamber and so the extension movement of the differential cylinder is faster than in control arrangements without differential circuit.

The DE 38 13 020 A1 shows a device for feed control of a double-acting hydraulic cylinder with means for setting a counter-pressure.

Furthermore, in the DE 195 24 900 A1 and in the JP 2003 185042 A pilot operated pressure relief valves disclosed.

In the US 3,160,076 a control arrangement for actuating the blade and the boom of a wheel loader, bulldozer or the like is disclosed. In this case, the control arrangement is designed with a pressure relief valve, via which the load pressure is limited to the two hydraulic cylinders. The pressure limiting valve is designed with a pressure switching stage, which makes it possible to limit the load pressure to a higher pressure when operating the blade alone, as it is the case when operating the boom or actuation of both hydraulic cylinders.

In the case of such control arrangements, under certain operating conditions, the action of external forces can lead to overloading and kinking of the piston rod. This is the case, for example, when the soil is to be removed while the blade is pivoted down and placed on the ground and the wheel loader then pulls off the ground in reverse. Runs the blade during this peeling on an obstacle, such as a boulder, so the blade holding the blade in the peel position of the differential cylinder is subjected to pressure and can bend.

In contrast, the invention has for its object to provide a hydraulic control arrangement and a pressure relief valve, by which damage to a differential cylinder of the control arrangement can be prevented.

This object is achieved with regard to the hydraulic control arrangement by the features of patent claim 1 and with respect to the pressure relief valve by the features of claim 14 or 15.

According to the invention, the hydraulic control arrangement is designed with a differential cylinder. Its pressure chambers can be connected via a control valve arrangement with a pump or a tank, so that a piston rod of the differential cylinder extends or retracts. The pressure in the effective pressure in the supporting pressure chamber is limited in the inventive solution via a pilot-operated pressure relief valve. Its pilot stage is designed with a pressure changeover stage, via which the pressure set at the lower pressure in the other pressure chamber of the pressure limiting valve is lowered to the extent that overloading of the piston rod is reliably prevented. In this case, a control surface of the pilot stage is acted upon by the pressure in the other pressure chamber, so that the limit pressure at which the pressure relief valve opens, in response to this pressure is variable. Such a solution is characterized by an extremely simple compact design with increased reliability.

According to the invention, it is particularly preferred if the differential cylinder can be controlled via the control valve arrangement in differential circuit, in which the annular space is connected to the cylinder space when the piston rod is extended.

The pressure switching stage preferably has a control spring of the pilot stage of the pressure relief valve acting on the clamping piston, which is acted upon in the direction of increasing the spring preload pressure in the piston rod side annulus and in the direction of lowering the spring preload pressure in the other, effective in the support pressure chamber (cylinder chamber), wherein in this Direction effective control surface of the tensioning piston is smaller than the effective in the direction of increasing the spring bias control surface.

The basic structure of a pressure relief valve used in the control arrangement according to the invention is per se from the DE 100 62 428 A1 the applicant known. Different from this solution is in a preferred embodiment, that the clamping piston of the pressure switching stage is acted upon in the direction of increasing the bias of a pilot valve cone acting on the control spring from a control pressure corresponding to the pressure in the other pressure chamber, which drops when exposed to an external force when he not already tank pressure. An effective in the direction of a reduction of the control spring bias smaller control surface is acted upon by the pressure in the effective in the support direction pressure chamber. In the known solution, however, the clamping piston of the pressure changeover stage is acted upon in the direction of increasing the bias of the pressure at the inlet of the pressure relief valve, which corresponds to the pressure in the effective pressure in the support direction. In the direction of discharge of the control spring, the tensioning piston is acted upon in the known solution by an external control pressure - this known pilot operated pressure relief valve could not be used without changes in the inventive solution.

In a variant of the embodiment with clamping piston is dispensed with the smaller control surface.

In most applications, the problem of overloading the piston rod described above will occur when it is almost completely extended, ie, in this case, the pressure chamber effective in the support direction of the bottom-side cylinder chamber during the another pressure space in which the pressure is lowered when exposed to an external load, the piston rod side annulus is.

The area ratio between the control surface of the tensioning piston and the pilot valve seat surface is <1.5 in one embodiment.

The control arrangement can be made particularly compact if a pilot piston of the pressure limiting valve is provided with a longitudinal channel, is passed over the control oil from a spring chamber of a main stage of the pressure relief valve to the smaller control surface.

In such a variant of the pilot piston is preferably carried out with a projection which dips sealing into a recess of the clamping piston. The end face of this recess then forms the smaller control surface, wherein the effective size of this surface is equal to the cross-sectional area of the projection.

In a particularly simple embodiment, the two control surfaces are formed on a pilot piston, wherein a smaller control surface from the pressure in the other pressure chamber (for example, piston rod side) and the larger control surface from the pressure in the other pressure chamber of the consumer (for example cylinder chamber) is acted upon - the tensioning piston then be waived. -

For maintenance purposes or the like, the pressure limiting valve is designed with an emergency opening, via which the input connection can be connected directly to the tank connection.

In a preferred embodiment, the control valve arrangement used in the control arrangement has a metering orifice formed by a continuously adjustable directional valve, which is followed by a LUDV pressure compensator. It is particularly preferred if the pressure medium supply takes place via a pump whose flow rate is adjustable in dependence on the highest load pressure of the entire system - the control arrangement then represents a LUDV system.

Other advantageous developments are the subject of further subclaims.

• Figur 1 einen Schaltplan einer erfindungsgemäßen hydraulischen Steueranordnung;
• Figur 2 einen Längsschnitt durch ein vorgesteuertes Druckbegrenzungsventil mit Spannkolben der Steueranordnung aus Figur 1;
• Figur 3 ein Schaltsymbol des Druckbegrenzungsventils aus Figur 2;
• Figur 4 einen Längsschnitt durch ein weiteres Druckbegrenzungsventil mit Spannkolben;
• Figur 5 ein Schaltsymbol dieses Druckbegrenzungsventils;
• Figur 6 einen Längsschnitt durch ein Ausführungsbeispiel eines Druckbegrenzungsventils ohne Spannkolben;
• Figur 7 ein Schaltsymbol dieses Ausführungsbeispiels und
• Figur 8 Kennlinien der in den Figuren 2, 4 und 6 dargestellten Druckbegrenzungsventile.

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

• FIG. 1 a circuit diagram of a hydraulic control arrangement according to the invention;
• FIG. 2 a longitudinal section through a pilot operated pressure relief valve with clamping piston of the control arrangement FIG. 1 ;
• FIG. 3 a switching symbol of the pressure relief valve off FIG. 2 ;
• FIG. 4 a longitudinal section through another pressure relief valve with clamping piston;
• FIG. 5 a switching symbol of this pressure relief valve;
• FIG. 6 a longitudinal section through an embodiment of a pressure relief valve without clamping piston;
• FIG. 7 a switching symbol of this embodiment and
• FIG. 8 Characteristic curves in the Figures 2 . 4 and 6 illustrated pressure relief valves.

In FIG. 1 is a circuit diagram of a directional control valve element 1 of a mobile control block, via which several consumers of a mobile implement, such as a wheel loader can be controlled. This in FIG. 1 illustrated directional control valve element 1 of the mobile control block is used to control an actuating cylinder 2, via which a blade mounted on a boom is pivotable.

The disk-type directional control valve element 1 has a pressure port P, a tank port T, two working ports A1, B1, and two control ports a1, b1, another control port x and a LS port LS. In the illustrated embodiment, the control block is designed as a LUDV system, via which a load pressure-independent flow distribution is possible. In such LUDV systems, a pump with variable delivery volume, for example, a variable displacement pump depending on the highest load pressure of the consumer is controlled.

The LUDV directional control valve element 1 has a continuously adjustable directional control valve 4 whose valve spool can be acted upon by a control pressure via the two control ports a1, b1 and is thus displaceable from a spring-biased central locking position into a multiplicity of control positions identified by (a) or (b). The directional control valve 4 has at least one pressure connection P, a tank connection T, two working connections A, B as well as two further connections D and D '. The directional control valve 4 forms a directional part, which is indicated by the two intersecting or branching arrows, and a speed part, which is formed by a variable metering orifice 5, which lies between the connections D and D '.

The two working ports A, B of the directional control valve 4 are connected via working lines, referred to below as feed line 6 and return line 8, to the working port A1 or to the working port B1. Between the working port B of the directional control valve 4 and the working port B1 a so-called low-leakage valve 10 is arranged in the return line 8, which consists in principle of a logic valve 12 and a pilot control valve 14. The logic valve has a stepped valve body which is loaded by a spring accommodated in a spring chamber in the closing direction. The spring chamber is connected via a throttle to the working port B1 of the directional control valve element. The pilot valve 14 is biased into a blocking position and can be switched by means of an actuating piston 16 from this blocking position into a passage position in which the spring chamber of the logic valve 12 is connected via a tank control channel 17 with a tank port T connected to the tank channel 18, so that the spring chamber of the logic valve 12 is depressurized. The graduated valve body of the logic valve 12 can thus be lifted at a pressure fluid flow in the return line 8 to the actuating cylinder 2 already due to a non-return function and at a flow of pressure medium from the actuating cylinder 2 to port B of the directional control valve with relief of the spring chamber of its valve seat. The actuating piston 16 is acted upon via a control branch passage 20 with the pressure at the control port a1, wherein due to a large area of the actuating piston 16, a comparatively large force is applied to the pilot control valve 14. Since the construction of such a low-leakage valve 10 is known, further relevant embodiments are unnecessary.

The two working ports A1, B1 of the directional control valve element 1 are via working lines 24, 26 with a bottom-side cylinder chamber 28 and a piston rod-side annular space 30 of the actuating cylinder 2 designed as a differential cylinder connected.

The directional control valve element 1 is further penetrated by a connected to the pressure port P pump channel 32. From this branches off an inlet channel 34, which leads to the connection D of the directional control valve 4. The port D 'of the directional control valve is connected via a connecting channel 36 to an input port P of a LUDV pressure compensator 38, the pressure compensator piston is acted upon in the opening direction by the pressure in the connecting channel 36 and in the closing direction by the force of a spring and the highest load pressure of the operated consumer , which is tapped via a connected to the LS port LS LS channel 40.

The pressure compensator is thus acted upon in the opening direction by the pressure downstream of the metering orifice 5. An output port A of the pressure compensator 38 is connected via a pressure compensator channel 42 and a check valve 44 to the input port P of the directional control valve 4. The tank connection T is connected to the tank channel 18 by means of a drainage channel 46.

The pressure in the return line 8 connected to the annular space 30 is limited by a secondary pressure limiting valve 48, which is arranged in a discharge channel 50 which branches off in the region of the pressure medium flow path between the logic valve 12 and the associated working port B1 of the return line 8 and with the tank channel 18 is connected. The pressure protection of the associated with the cylinder chamber 28 feed line 6 via a pilot-operated pressure relief valve 52 which is arranged in a likewise connected to the tank channel 18 channel 54, in the area branches off from the flow line 6 between the directional control valve 4 and the working port A1.

The pilot-operated pressure limiting valve 52 and the pressure limiting valve 48 are each designed with a Nachsaugfunktion, so that pressure medium can be sucked from the tank channel 18 to avoid cavitation in a pulling load.

The pilot-operated pressure relief valve 52 is, as described below in more detail with reference to Figures 2 and 3 is explained, from a main stage, a pilot stage and a pressure changeover stage 56. The latter makes it possible to change the set pressure on the pilot operated pressure relief valve 52. This schematic in FIG. 1 illustrated pressure changeover stage 56 has a clamping piston 58, on which a control spring 60 of the pilot stage is supported. A larger control surface of the clamping piston 58 is acted upon by the pressure in a pilot channel 62, which leads to the control terminal X of the directional control valve element 1, which in turn is connected via a line 64 with the leading to the annulus 30 working line 26. On a comparatively small control surface of the clamping piston 58, the pressure acts in the flow line 6, which is tapped off via the channel 54 and via a tapping channel 66.

For extending a piston rod 68, the directional control valve 4 is brought into one of its (a) marked positions by the control port a1 is acted upon by a control pressure. This control pressure can be adjusted for example via pressure reducing valves, which reduce the pressure in a control circuit to a suitable control pressure.

The pressure medium then flows from the variable displacement pump via a pump line, not shown, to the pressure port P and from there via the pump channel 32, the inlet channel 34 to port D of the directional control valve, from there via the metering orifice 5 corresponding to the control pressure to the port D 'of the directional control valve 4 and via the connecting channel 36 to the connection P of the LUDV pressure compensator 38. This LUDV pressure compensator 38 arranged downstream of the metering orifice 5 throttles the pressure medium volume flow so strongly that the pressure after all metering orifices of the system is the same and preferably corresponds to the highest load pressure or slightly above this is. D. h., With a shortage of several consumers changes to the pressure downstream of the orifices nothing. The pump pressure is applied in the same way to all metering orifices of the system, so that the pressure difference at all metering orifices alters in the same way if the pump pressure decreases in the event of an undersupply - the flow distribution between the metering orifices is maintained (load pressure-independent flow distribution).

The thus throttled pressure medium flow then flows through the pressure balance channel 42, the input port P and the working port A of the directional control valve 4 and the flow line 6 and the working line 24 to the cylinder chamber 28. The piston rod 68 extends, wherein the pressure medium displaced from the annular space 30 via the working line 26 and the working port B1 flows. By the control pressure at the control port a1, the pilot control valve 14 is brought from its spring-biased locking position to its passage position, so that the spring chamber of the logic valve 12 relieved and this is opened by the pressure in the drain line 8, so that the pressure fluid continues to the working port B of the directional control valve 4th flows and there to that of the pump subsidized pressure medium flow is summed. Tank connection T is shut off in positions (a). The pilot operated pressure relief valve 52 remains set to a comparatively high pressure, which should be, for example, 380 bar. As will be explained in more detail below, this higher pressure is adjusted by acting on the larger control surface of the clamping piston 58, the pressure in the annular space 30, which is at least as large as the pressure in the cylinder chamber 28, which is the smaller control surface of the differential circuit Tensioning piston 58 acted upon.

To retract the piston rod 68, the directional control valve 4 is displaced by applying a control pressure to the control port b1 in one of its (b) marked positions, in which case the cylinder chamber 28 is connected to the tank channel 18 and the annular space 30 with the pump channel 32, so that Pressure medium is conveyed into the annular space 30 and flows back from the cylinder chamber 28 displaced pressure fluid to the tank T.

It is now assumed that the removal of a soil initially described is to take place. As mentioned above, the piston rod 68 is extended for this purpose (directional control valve in position (a)) and thus the blade is completely pivoted and then the directional control valve is returned to its spring-biased center position. The bucket then rests on the ground and the wheel loader reverses to pull off the ground. When the blade encounters an obstacle, the piston rod 68 is subjected to pressure in the direction of retraction, as a result of which the pressure in the annular space 30 and correspondingly the pressure in the control channel 62 drop. By this decrease in pressure in the annular space 30 of the clamping piston 58 is under the action of the control spring 60 and the smaller control surface acting pressure in the cylinder chamber 28 in the direction of relief of the control spring 60 moves. The tensioning piston 58 is moved to the rear against a stop and the control spring 60 is relieved, so that the pressure relief valve is set to a much lower pressure, for example, 100 bar. When this pressure in the cylinder chamber 28 is exceeded, the pilot-operated pressure relief valve 52 opens, so that damage to the piston rod 68 is prevented by excessive pressure load.

The used pilot-operated pressure relief valve 52 will be described below with reference to Figures 2 and 3 explained.

FIG. 2 As already mentioned, this has a main stage 70, a pilot stage 72 and the pressure changeover stage 56. The basic structure of the main stage 70 and the pilot stage 72 is essentially made of DE 100 62 426 A1 known, so that only the components required for understanding the invention are described and reference is otherwise made to this pre-published document. The pilot-operated pressure relief valve 52 is designed in cartridge construction and has a housing 74, on which an end-side pressure port P and a radial, formed for example by a bore star tank port T is formed. In the housing 74, a valve slide 76 designed with a sliding seat is guided in a valve bore 78, which is prestressed against a seat edge 82 via a weak compression spring 80. In the illustrated closed position, the connection between the input port P and tank port T is shut off. The valve spool 76 is hollow, wherein in an axially projecting end face a nozzle bore 84 is formed, which extends inwardly toward a spring chamber 110 extended for the compression spring 80. At the in FIG. 2 arranged on the right rear side of the valve spool 76 is a radial collar 86 is formed. This forms a stop for a Nachsaugring 88 which is sealingly guided in an annular space between a radially enlarged portion of the valve bore 78 and the outer periphery of the valve spool 76. In the FIG. 2 left end face of the Nachsaugrings 88, is acted upon via a throttle gap 90 with the pressure at the tank port T.

A sealing edge 92 is formed in the radially widened region of the valve bore 78, against which a seat body 94 inserted in a further enlarged region of the valve bore 78 rests. This is biased by means of a screwed into the housing 70 pilot housing 96 against the sealing edge 92. On the seat body 94, a pilot valve seat 98 is formed, against which a pilot valve cone 100 is biased by the control spring 60. For axial guidance of the pilot valve cone 100 has a collar 102, the outer circumference is guided in a provided with two longitudinal grooves guide bore 104 of the seat body 94. At the in FIG. 2 Left end face of the seat body 94 is formed an axial projection in which a closed blind hole 106 is provided to the left, which widens to the pilot valve seat 98 and which is connected via radial bores 108 with the spring chamber 110 for the compression spring 80. The in FIG. 2 the space 93 formed on the right of the seat 92 is connected to the tank connection T via a sloping channel 95. This space 93 is also connected via connecting holes 97 with the space encompassed by the seat body 94. About the longitudinal grooves in the guide bore 104 of the seat body 94 and the control spring 60 receiving space is connected to the tank.

From the collar 102 of the pilot valve cone 100 to the right extends a projection 111 whose end portion dips into a recess 112 of the clamping piston 58, which is axially displaceably guided in a through hole 114 of the pilot housing 96. This through-bore 114 extends coaxially with the valve bore 78. The pilot-control cone 100 and its projection 111 are penetrated by a longitudinal channel 116 which opens into the control chamber 118 defined by the recess 112 and the projection 111. D. h., Via the longitudinal channel 116 and the radial bores 108 of the pressure applied in the spring chamber 110 pressure is tapped and acts on a comparatively small control surface 120 which is formed by the end face of the recess 112.

The control spring 60 is on the in FIG. 2 supported on the left end face of the clamping piston 58, so that this rests in its illustrated basic position on a screwed into the through hole 114 stop screw 122. The through hole 114 opens at the right end face of the pilot housing and forms a port X1 of the pilot operated pressure relief valve 52, to which the in FIG. 1 shown control channel 62 is connected. The stop screw 122 is annular, so that the pressure at the control terminal X1 also acts on the back of the clamping piston 58, which forms a substantially larger compared to the control surface 120 control surface 124.

At a distance to the left of the tensioning piston 58, a radial shoulder acting as a stop 126 is formed on the through-bore 114, which moves the axial travel of the tensioning piston 58 to the left (FIG. FIG. 2 ) limited.

The symbol of the in FIG. 2 illustrated pressure relief valve 52 is in FIG. 3 strongly schematized shown. Shown are the main stage 70, the pilot stage 72 and the pressure switching stage 56 with the clamping piston 58 and the pilot housing 96. The larger control surface 124 of the clamping piston 58 is acted upon by the pressure in the control channel 62 and the smaller control surface 120 with pressure at the input terminal P over the longitudinal channel 116 and the spring chamber 110 and the nozzle bore 84 is tapped (see FIG. 2 ). In FIG. 3 only the reference numeral for the longitudinal channel 116 is shown.

The clamping piston 58 acts on the control spring 60, which acts on the valve spool 76 of the main stage 72 in the closing direction. In the opening direction acts on the valve spool 76, the pressure at the input port P, which also rests in the channel 54 and in the flow line 6.

For maintenance purposes, the pressure port P of the pressure relief valve 52 can be connected by hand to the tank port T. This is in FIG. 3 indicated with the manually operable switching valve 128. When switching this switching valve 128 in its passage position of the input port P of the pressure relief valve 52 is relieved to the tank channel 18 out. At the in FIG. 2 shown concrete Ausfürungsbeispiel this emergency opening is formed by the interaction of the seat body 94 with the sealing edge 92.

When fully screwed pilot housing 96 of the seat body 94 sits firmly on the sealing edge 92 - this corresponds to the closed position of the switching valve 128 (see FIG. 3 ). For the emergency opening, the manually accessible pilot housing 96 is unscrewed somewhat out of the housing 74, so that the seat body 94 lifts off from the sealing edge 92 and the spring chamber 110, in which the pressure is usually applied to the input terminal P. the slant channel 95 is connected to the tank port T or more precisely the tank channel 18 - the valve spool 76 can then be moved by the pressure at the input port P against the force of the comparatively weak compression spring 80 to the right, so that the connection to the tank port T is opened.

In the case in which, for example, a pulling load - can not be promoted by the pump enough pressure fluid to the cylinder chamber 28 and thus the corresponding load pressure drops below the tank pressure, the Nachsaugring 88 is moved by the higher tank pressure to the right and runs on the Radial collar 86, so that the valve spool 76 is taken and the connection from the tank port T to the input port P is opened, so that pressure fluid can be sucked from the tank.

As already on the basis of FIG. 1 explained acts in normal operation, for example when pivoting or tilting of the blade, ie, when extending the piston rod 68 at the control port X1 of the pressure relief valve 52, the pressure in the annulus 30, at least as high in the differential circuit (control positions (a) of the directional control valve 4) as the pressure in the cylinder chamber 28 is. That is, the pressure acting on the larger control surface 124 is at least equal to the pressure acting on the smaller control surface 120 corresponding to the pressure in the cylinder space 28. The force acting in one direction on the tensioning piston is the sum of the force of the control spring 60 plus the pressure force generated on the control surface 120, which is equal to the cross-sectional area of the projection 111 within the recess 112, by the pressure prevailing in the spring space 110 becomes. The spring force, in turn, is equal to a compressive force generated by the boundary pressure on a surface containing the differential area between the cross-sectional area of the pilot valve cone 100 at the seat 98 and the control surface 120 is. Finally, the force acting on the tensioning piston in the one direction when the higher limit pressure is reached corresponds to a compressive force produced by the higher limit pressure on the cross-sectional area of the pilot valve cone on the seat 98.

In the opposite direction acts on the clamping piston 58, a pressure force which is generated by the pressure applied to the control surface 124 in connection X1.

Assuming the area ratios off FIG. 2 can be adjusted with a certain pressure in the terminal X1 an approximately six times higher limit pressure, the highest limit pressure is given by contact of the clamping piston on the stop 126 and the then existing bias of the control spring 60. When the specified pressure at port X1 is reached, the tensioning piston 58 is moved from the position in FIG FIG. 2 shifted to the left until it runs onto the stop 126. As a result, the control spring 60 is tensioned - at the pilot operated pressure relief valve 52, the higher pressure is set. For example, to set 390 bar, 65 bar in connection X1 is sufficient. When removing the bottom and an external load, which acts on the piston rod 68 in the retraction, the pressure in the annular space 30 decreases, if he was not already tank pressure, while the pressure in the cylinder chamber 28 increases. The geometry of the clamping piston 58 is selected so that from a certain pressure difference between the pressure chambers 28, 30 of the clamping piston 58 lifts by the relief of the control surface 124 from the stop 126 and is moved against the stop screw 122. This return movement is assisted by the pressure acting on the smaller control surface 120 - the bias of the control spring 60 is reduced and corresponding to the release pressure of pilot operated pressure relief valve 52 set to a lower pressure (100 bar). This pressure is chosen so that damage to the piston rod 68 can be reliably avoided. Below a certain pressure in the terminal X1 of the clamping piston 58 is also at tank pressure in the cylinder chamber 28 on the stop screw 122, namely, when the pressure force is less than the force of the relaxed spring 60.

The small control surface 120 of the clamping piston 58 causes, when the pressure limiting valve 52 responds, the tensioning piston 58 is acted upon in the direction of relaxing the control spring 60 by a force which is as great as the force generated by the inlet pressure (P) on the entire seating surface of the pilot valve seat 98 is. For the control spring 60, however, only the differential area between the valve seat surface and the small control surface 120 is relevant, so that the pressure in the annular space 30 of the actuating cylinder 2 must drop relatively sharply, so that the pressure relief valve responds. In a pressure relief valve 52 with the in FIG. 2 shown geometry ratios would be sufficient at an assumed limit pressure of for example 360 bar, a pressure in the annular space 30 and corresponding to the larger control surface 124 acting pressure of about 60 bar to hold the clamping piston 58 to the stop 126. Only when falling below this 60 bar opens the pressure relief valve 52, then at a pressure of about 20 bar, the clamping piston 58 runs onto the stop screw 122 and thus determines the lower limit pressure, which would then be about 120 bar.

Based FIG. 4 an embodiment is explained in which the pressure relief valve 52 already responds at a much smaller pressure drop in the annular space 30 of the actuating cylinder 2. This will be at the in FIG. 4 illustrated embodiment essentially achieved in that the additional smaller control surface 120 is omitted and the area ratio between the effective diameter of the clamping piston 58 and the pilot valve seat diameter is chosen substantially smaller than in the above-described embodiment. This area ratio is at the in FIG. 4 illustrated embodiment, about 1.12, ie, the pilot valve seat area A ₂ is 1.12 times greater than the effective area A _{1 of} the clamping piston 58th

The basic structure of in FIG. 4 illustrated embodiment corresponds to that of FIG. 2 , Accordingly, this is also in FIG. 4 illustrated embodiment with a main stage 70, a pilot stage 72 and a pressure switching stage 56 executed. The main stage 70 with the valve spool 76, the pressure spring 80 arranged in the spring chamber 110, the cartridge-shaped housing 74 and the Nachsaugring 88 corresponds to the main stage 70 of the above-described embodiment, so that the simplicity is referred to the relevant embodiments. The pilot stage 72 and the switching stage 56 are substantially integrated into the pilot housing 96, which is screwed into the cartridge-shaped housing 74 and presses the seat body 94 against the sealing edge 92 (in the illustrated basic position). Similar to the above-described embodiment, the sealing body 94 is designed with an axial projection 130, in which the blind hole 106 is formed, which opens via the radial bores 108 in the spring chamber 110. In the blind hole 106 of the seat body 94, a damping piston 132 is guided axially displaceable, which via damping gaps (in FIG. 4 not shown in detail) a pressure medium connection in the direction of the pilot valve seat 98 allows.

Against this pilot valve seat 98, a spherical pilot valve body is biased in this embodiment, which is also referred to as a pilot valve cone 100 for simplicity. This is supported by a mushroom-shaped spring plate 134 on which the control spring 60 engages, which in turn is supported by a further spring plate 136 on the clamping piston 58. The outer circumference of the mushroom-shaped spring plate 134 is guided within the seat body 94. The control spring 60 receiving space 93 is - as in the above embodiment - connected to the tank port T.

In the illustrated basic position of the clamping piston 58 rests with a radially enlarged stopper head 138 on the screwed into the pilot housing 96 stop screw 122, so that the clamping piston 58 in the direction of increasing the bias of the control spring 60 by the pressure at the control port X1 (pressure in the annular space 30th ) is acted upon. The clamping piston 58 is guided along a through hole 114 of the pilot housing 96 as in the embodiment described above. This through hole 114 widens to the right (view after FIG. 4 ) to the terminal X1, wherein on an annular shoulder a stopper piece 140 is supported, which corresponds to the stop 126 and thus the axial stroke of the clamping piston 58 to the left in FIG. 4 limited. The acted upon by the pressure at the control terminal X1 effective control surface A ₁ is defined by the outer diameter of the radially recessed portion of the clamping piston. In the above extension of the piston rod 68 (in differential circuit) thus acts on the resulting effective area A _{1 of} the clamping piston 58, the pressure in the annular space 30 so that the clamping piston 58 is moved at sufficient pressure in the annular space 30 from its illustrated home position to the left until the stopper head 138 on the stopper piece 140th runs up and the control spring 60 is stretched - the upper limit pressure is set.

The pilot stage 72 opens when the effective pressure at the pilot valve seat 98 is sufficient to lift the pilot valve cone 100 from the pilot valve seat 98. In the opening direction acts on the pilot valve seat 98 with the cross-sectional area A _2, the pressure at the pressure port P of the nozzle bore 84, the spring chamber 110, the radial bores 108 and limited by the damping oil 106 damping gap is tapped. In the illustrated embodiment, the area ratio A ₁ / A _{2 is} relatively small (for example, 1.12) executed, so that even at a much higher pressure in the annulus 30 than in the above embodiment, the pilot stage 72 is open. Assuming, for example, that the limiting pressure is 380 bar, the pressure relief valve would correspondingly at a pressure of about 340 bar - ie much earlier than in the FIG. 2 illustrated embodiment - open. This early opening is further supported by the fact that in FIG. 4 illustrated embodiment, an effective in the direction of relaxation of the control spring 60 control surface (120 in FIG. 2 ) is missing. Decreases in this embodiment, the pressure in the annular space 30 further, for example, to 110 bar, so the stopper head 138 comes into contact with the stop screw 122, so that the lower limit pressure (minimum bias of the control spring 60) is set. This minimum limit pressure corresponds to in the embodiment according to FIG. 4 corresponding to the area ratio A _{1 /} A ₂ then about 123 bar. In the intervening areas, ie, at pressures in the annular space 30 between 110 and 340 bar, the boundary pressure increases linearly in accordance with this area ratio.

The symbol of the in FIG. 4 illustrated embodiment is in FIG. 5 shown. This symbol is essentially the same as that FIG. 3 wherein the pressure switching stage 56 has no effective in the direction of relaxation of the control spring 60 control surface 120. The illustrated control oil nozzle is as in the embodiment according to the Figures 2 and 3 formed by the nozzle bore 84.

FIG. 6 shows a further simplified embodiment of a pressure relief valve according to the invention, in which dispensed with the use of a clamping piston.

The basic structure of the valve is identical to that with reference to the guide and the structure of the pilot valve cone 100 FIG. 2 described embodiment, so that in terms of the description of the main stage 70 with the valve spool 76, the compression spring 80 and the Nachsaugring 88 and with respect to the seat body 94 and screwed into the housing 74 of the main stage 70 pilot housing 96 to the comments on FIG. 2 is referenced. The outer contour of the pilot valve cone 100 also corresponds to the in FIG. 2 illustrated embodiment, ie, to the guided in the seat body 94 collar 102 connects to the right to a non-hollow projection 111, the cylindrical end portion 142 passes through a guide portion 144 of the pilot housing 96 which is formed by a radially recessed portion of the through hole 114. In the FIG. 6 Right end face 146 of the pilot valve cone 100 defines a control chamber 148, which is acted upon by the control oil pressure at the control terminal X1. As in the embodiment according to FIG. 2 the pilot valve cone 100 is biased by the control spring 60 against the pilot valve seat 98, the effective area in the Representation according to FIG. 6 is designated with the cross-sectional area A ₂ , while the active surface of the end face 146 in FIG. 6 marked with A ₁ . The control spring 60 is supported on a fixed annular end face 150 of the pilot housing 96.

In this embodiment, the two limit pressures are determined by the ratio of the areas A _{1 /} A ₂ . In the case in which the pressure in the annular space 30 is approximately zero, the pressure at the control port X1 and thus also the pressure in the control chamber 148 is approximately zero, so that the end face 146 is not subjected to a control oil pressure - the pilot valve cone 100 becomes thus biased solely by the force of the control spring 60 against its pilot valve seat 98, so that the lower limit pressure is set. If, in a differential circuit, the pressure in the annular space 30 is substantially equal to the pressure in the cylinder chamber 28 of the actuating cylinder 2, both the end face 146 and the limited by the pilot valve seat 98 face area of the pilot valve piston 100 are subjected to the same pressure, so that this pressure at the Area difference A ₂ -A _{1 is} effective and the upper limit pressure is set.

The symbol of the in FIG. 6 shown pressure relief valve 52 is in FIG. 7 shown. Accordingly, in this embodiment, the bias of the control spring 60 is not changed, but only the pressure acting on the pilot control in the opening and closing direction, a change in the pressure at the control terminal X1 always results in a change of the set limit pressure. If this limiting pressure p _G set at the pressure limiting valve 52 is applied as a function of the pressure p _{X1 present} at the control port _X1 (pressure in the annular space 30), the result is in FIG. 8 continuous line characteristic. Accordingly, in an embodiment according to FIG FIG. 6 (solid line in FIG. 8 ), the areas in which a change in the pressure p _{X1 has} no influence on the set limit pressure p _G , while in the embodiments according to the Figures 2 and 4 the upper and lower limit pressures are characterized by the horizontally extending portions in which a change in the control pressure PX1 has no influence (dot-dashed line). The intermediate linear increase depends essentially on the above-described area difference of the effective control surfaces.

Disclosed are a hydraulic control arrangement and a dedicated pilot operated pressure relief valve. The hydraulic control assembly has a differential cylinder with a piston rod side pressure chamber and a bottom pressure chamber, which are connectable via a control valve assembly for actuating the differential cylinder with a pump or a tank. The pressure in a pressure chamber is limited by a pilot-operated pressure relief valve, which is designed with a pressure changeover stage, via which the pressure set on the pressure limiting valve can be lowered in dependence on the pressure in the other of the pressure chambers.

LIST OF REFERENCE NUMBERS

1

Way valve element

2

actuating cylinder

4

way valve

5

metering orifice

6

supply line

8th

Return line

10

Low-leak valve

12

logic valve

14

Pilot switching valve

16

actuating piston

17

Tank control channel

18

tank channel

20

Control branch channel

24

working line

26

working line

28

cylinder space

30

annulus

32

pump channel

34

inlet channel

36

connecting channel

38

LUDV pressure

40

LS-channel

42

Pressure balance channel

44

check valve

46

drain channel

48

Pressure relief valve

50

relief channel

52

pilot operated pressure relief valve

54

channel

56

Druckumschaltstufe

58

clamping piston

60

control spring

62

control channel

64

management

66

Abgreifkanal

68

piston rod

70

main stage

72

pilot stage

74

casing

76

valve slide

78

valve bore

80

compression spring

82

seat edge

84

nozzle bore

86

radial collar

88

anti-cavitation

90

throttle gap

92

sealing edge

93

room

94

seat body

95

oblique channel

96

pilot housing

97

connecting bore

98

Pilot valve seat

100

Pilot valve cone

102

Federation

104

guide bore

106

Blind hole

108

radial bore

110

spring chamber

111

head Start

112

recess

114

Through Hole

116

longitudinal channel

118

control room

120

small control surface

122

stop screw

124

larger control surface

126

attack

128

switching valve

130

axial projection

132

Dämpfungskölbchen

134

spring plate

136

another spring plate

138

stop head

140

stop piece

142

end

144

guide section

146

face

148

control room

150

Annular face

Claims (16)

1. A hydraulic control arrangement comprising a differential cylinder (2) having a first pressure chamber (28) and a second pressure chamber (30) which can be connected to a pump or a tank by means of a control valve arrangement (4, 38) in order to actuate the differential cylinder (2), wherein the control valve arrangement comprises a continuously variable directional control valve (4) forming a variable metering orifice (5), characterized by a pilot-operated pressure relief valve (52) arranged on an advance line (6) between the continuously variable directional control valve (4) and the differential cylinder (2) for limiting the pressure in one of the pressure chambers (28), wherein a control surface (124, 146; A₁) to which the pressure prevailing in the other one of the pressure chambers (30) is applied is provided in a pilot stage (72) of the pressure relief valve (52), and wherein the output of the pilot-operated pressure relief valve (52) is connected to a tank passage (T).
2. The hydraulic control arrangement according to claim 1, wherein the first-mentioned pressure chamber is a cylinder chamber (28) on the bottom side and the other pressure chamber is an annular chamber (30) on the piston rod side.
3. The hydraulic control arrangement according to claim 2, wherein the two pressure chambers (28, 30) can be connected to each other by means of the control valve arrangement (4) in order to actuate the differential cylinder (2) in the differential circuit.
4. The hydraulic control arrangement according to any one of the preceding claims, wherein the pilot stage comprises a tensioning piston (58) pressurized by a control spring (60) to which the pressure prevailing in the other pressure chamber (30) is applied in the direction of increase in the spring bias.
5. The hydraulic control arrangement according to claim 4, wherein the pressure prevailing in the first-mentioned pressure chamber (28) is applied to a comparatively smaller control surface (120) of the tensioning piston (58) in the direction of relief of the control spring (60).
6. The hydraulic control arrangement according to claim 5, wherein the surface ratio (A₁/A₂) between the control surface (124, 146) of the tensioning piston (58) and the active surface of the pilot valve seat (98) is less than 4, preferably less than 1.5.
7. The hydraulic control arrangement according to claim 5 or 6, wherein a pilot piston (100) of the pressure relief valve (52) is provided with a longitudinal passage (116) through which control oil is guided from a spring chamber 110) of a main stage (70) to the smaller control surface (120).
8. The hydraulic control arrangement according to claim 7, wherein the pilot piston (100) has a projection (111) which immerses into a recess (112) of the tensioning piston (58) in a sealing manner, the end face of the recess forming the smaller control surface (120).
9. The hydraulic control arrangement according to any one of the claims 1 to 3, wherein the control surface (146) is formed at a projection (111) of a pilot piston (100) such that the pressure prevailing in the other pressure chamber (30) acts upon the pilot piston (100) in the closing direction.
10. The hydraulic control arrangement according to claim 9, wherein the projection (111) passes through a spring chamber (93) of the control stage and immerses into a control chamber (148) to which the pressure prevailing in the other pressure chamber (30) is applied.
11. The hydraulic control arrangement according to any one of the preceding claims, wherein the pilot-operated pressure relief valve (52) includes a manually operable emergency opening.
12. The hydraulic control arrangement according to any one of the preceding claims, wherein downstream from the variable metering orifice (5) a pressure regulator (38) forming part of the hydraulic control arrangement is connected to which the pressure downstream of the metering orifice (5) is applied in the opening direction and the maximum load pressure is applied in the closing direction.
13. The hydraulic control arrangement according to claim 12, comprising a pump the delivery rate of which is adjustable in response to the maximum load pressure so that a pump pressure is above the maximum load pressure by a particular pressure difference.
14. A pressure relief valve for a control arrangement in accordance with any one of the preceding claims, comprising an inlet terminal (P) and an outlet terminal (T) and comprising a main stage (70), a pilot stage (72) including a tensioning piston (58) which acts on a control spring (60) of a pilot valve cone (100), characterized in that the tensioning piston (58) has a larger control surface (124) to which a control pressure can be applied via a control terminal (X1) in the direction of increase in the spring bias, and wherein the pressure prevailing at the inlet terminal (P) is applied to a smaller control surface (120) of the tensioning piston (58) in the direction of reduction of the control spring bias.
15. A pressure relief valve for a control arrangement according to any one of the preceding claims 1 to 13, comprising an inlet terminal (P) and an outlet terminal (T) and comprising a main stage (70) and a pilot stage (72), characterized in that a spring-biased pilot valve cone (100) of the pilot stage (72) has a control surface A₁ to which a control pressure (X1) acting in the closing direction is applied, and a control surface A₂ active in the opening direction to which the pressure prevailing at the inlet terminal (P) is applied and which is larger than the control surface A₁ acted upon in the closing direction.
16. The pressure relief valve according to claim 15, wherein the surface ratio A₁/A₂ of the control surface is ≤ 4, preferably 1 < A₁/A₂ < 1.5.

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