DE10063153A1 - Pressure relief valve for excavator has slide which moves against force of spring whose opposite end rests against transmission piston with chamber pressurised by fluid from inlet to move piston and increase compression of spring - Google Patents

Abstract

The pressure relief valve has a slide (8) which moves against the force of a spring (14). The opposite end of the springs rests against a transmission piston (16) with a transmission chamber (50) which is pressurised by fluid from the inlet (P) and moves the piston to increase the compression of the spring. The chamber (18) in which the spring is mounted is connected by a control channel (Y) to the valve outlet (T).

Description

The invention relates to a directly controlled pressure relief valve til according to the preamble of claim 1.

Such pressure relief valves are used for example for Absi Protection of hydraulic circuits for chassis and slewing gear drives used, either as a closed or open Hydraulic circuit are executable. For example when swiveling an uppercarriage of an excavator or when accelerating over the drive the hydraulic motor is driven by a high hydraulic pressure strikes so that these movements are initiated very quickly and the structures of a considerable acceleration with correspondingly high Mass moments of inertia are exposed. This sudden strain can damage the system. One is therefore be strives to make speed changes such that soft Transitions between the driving / turning states are made. there are the currently ar when a pressure medium volume jump occurs processing consumers, for example the slewing gear or the chassis damped accelerated or decelerated (shockless function). At derar term circuits, for example, is formulated in two opposite directions pressure relief valves the high pressure side with the Low pressure side of the hydraulic system connected, so that at one sudden pressure increase in the high pressure branch pressure peaks by opening the connection to the low pressure side can be removed.

DE 31 07 775 A1 discloses a pressure relief valve, that can be used in such circuits. In this known Solution, the valve body is designed as a cone and via a pressure spring biased into a closed position. The compression spring is on one Supported translation piston, its translation space via a nozzle is pressurized at the inlet port. Because of the pressure The translator piston can be set up in the translator room move to the valve cone so that the preload of the compression spring is changeable.  

When the pressure at the inlet port rises rapidly the valve cone is first lifted from its seat, so that the connection is open to the output connection. The pressure increase at the entrance Connection is reported through the nozzle in the translator room, and in egg ne corresponding axial movement of the booster piston implemented. because by the bias of the compression spring is increased, so that the Ventilke gel with a greater force towards its valve seat is tense. The pressure at the inlet port of the pressure relief valve increases accordingly and then reaches its maximum value; if the Booster piston runs onto an axial stop and thus the Sy stem pressure is limited to the maximum set operating pressure. That means through the time-dependent change in the compression spring preload The pressure build-up in the system is delayed, so that shock loads are dampened.

The filling of the rear translator room essentially depends on the speed at which the control oil passes over the injected nozzle. This volume flow is strongly dependent on viscosity, so that for example at low temperatures or a high Pressure at the inlet port a slower pressure build-up than at high Operating temperatures or low pressures results.

Another problem is that it is removed from the valve seat benem valve cone can occur that this due to the flow the pressure medium is excited to vibrate, which is in the spring Propagate the space of the main cone so that the booster piston is subjected to the vibrations that the opening behavior of the Influence pressure relief valve negatively.

In contrast, the invention is based on the object; a direct to create controlled pressure relief valve in which the response behavior is also improved at different system pressures.

This task is accomplished through a directly controlled pressure limitation solved valve with the features of claim 1.  

According to the invention, the directly controlled pressure relief valve is executed with a valve spool, which in its via a compression spring Closing position is biased. The one from the valve spool and a booster piston on the other hand axially limited spring space is over a control channel with the pressure at the outlet connection (tank pressure, never derdruck) acted upon, this control pressure at a distance from the off gear connection is tapped. By tapping into this at a distance the output connection controlled by the valve slide can be the Influence of the pressure limitation in the area of the opening cross section Minimize the flow forces occurring, so that these Cannot propagate pressure fluctuations into the spring chamber. The means that the spring chamber of the pressure relief valve is hydraulically white test separated from the pressure medium flow area, so that a results in much better responsiveness than with solutions where de the control oil is tapped in the pressure medium flow area.

The response can be further improved if instead a current regulator is switched on in the nozzle upstream of a translator room is set, through which a viscosity-independent filling of the translation is possible. By using this current regulator. minimize the influence of temperature and pressure fluctuations, so that the delay in the build-up of pressure even with different operations conditions is constant. The applicant reserves the right to limit Va riante of a direct operated pressure relief valve with valve body per (slide or piston), booster piston and integrated Stromre gel valve to make its own independent claim.

In a preferred embodiment is a pressure compensator column ben of the current controller in an axial bore of the booster piston leads so that a nozzle by axial displacement of the pressure compensator piston in the jacket of the translator piston is controllable. In the ground a further nozzle is formed in the pressure compensator piston. At this Variant it is important that the adjustable nozzle in the jacket of the over set piston with a relatively small diameter, so that the desired time delay is set.  

It is particularly advantageous if the booster piston on one Spindle is supported, the contact surface of the spindle spherical is led. This spherical or spherical design of the end face of the booster piston enables the balancing of manufacturing tolerances zen so that the booster piston does not tilt in the contact position is performed and is therefore torque-free. It is particularly advantageous if the radius of curvature of the spindle face is chosen so that the booster piston in the radially inner area on the spindle is applied.

The axial length of the valve arrangement is particularly small if the Valve slide on a guide projection of the booster piston allowed gert is, in the valve spool end portion of the translator piston a nozzle can be provided, which is used to prevent that fluctuations at the inlet port of the pressure relief valve are not correct dampens get to the input of the current controller.

Other advantageous developments of the invention are counter stood further subclaims.

The following is a preferred embodiment of the invention tion explained in more detail using schematic drawings.

Show it:

Fig. 1 a cross section through an inventive directly controlled pressure relief valve;

FIG. 2 shows a detailed illustration of an integrated current regulator of the pressure limiting valve from FIG. 2;

Fig. 3 is a dynamic characteristic of the pressure relief valve according to the invention and

Fig. 4 is a circuit diagram of an application example of two inventive pressure relief valves.

The pressure relief valve 1 shown in longitudinal section in FIG. 1 has a cartridge-shaped housing 2 , in which an axial input connection P and an output connection T formed by a bore star 24 are provided (tank or low-pressure connection). The housing 2 is screwed into a receiving bore 23 of a valve block 4 .

The housing 2 has a widening from the input port P to the right end portion in FIG. 1 towards the valve bore 6 , in which a valve slide 8 is guided. The valve designed as a seat slide slider 8 is biased with a seating cone 10 against a shoulder formed by a radial bore of the valve 6 the valve seat 12th This pretensioning takes place by means of a compression spring 14 which acts on a rear annular shoulder of the valve slide 8 . The Druckfe 14 is supported on a booster piston 16 which is also axially displaceably mounted in the valve bore 6 .

The front side of the valve spool 10 on the one hand and piston 16 on the other hand limited spring space 18 is connected via a man telbohrungsstern 20 with an annular space 22 which is formed by a receiving bore 23 of the valve block 4 and which extends to the area in which the tank connection T of the housing 2 opens. This jacket hole star 20 forms a Steueran circuit Y from. The axial distance between the casing bore star 20 and the opening valve 24 controlled by the valve spool 10 is so selected that the pressure in the area of the casing bore star 20 is not influenced by the flow forces which occur when the pressure limiting valve 1 flows from P to T.

The valve slide 8 is approximately cup-shaped, wherein the seat cone 10 is penetrated by a nozzle bore 26 which opens into an interior 28 of the valve slide 10 .

The booster piston 16 has a guide projection 30 , which extends in the axial direction into the interior 28 of the valve slide 8 and is penetrated by an axial bore 32 . A damping nozzle 34 is inserted into the valve slide-side end section of the guide projection 30 .

The guide projection 30 merges via a radial shoulder 36 into a basic body 38 guided along the inner circumferential wall of the valve bore 6 . The valve spool 8 in its contact position in front of exciting compression spring 14 is supported on the radial shoulder 36 . The rear end face 40 of the base body 38, which is removed from the radial shoulder 36 , is provided with a convexly curved, approximately spherical annular surface which merges into an axially projecting end section 42 . In this end section, a screw plug 44 is inserted, via which the axial bore 32 is closed on the end face.

In a radially enlarged part of the axial bore 32 , a flow control valve 46 is also added, which is described in more detail below with reference to FIG. 2. In the dargestell th in Fig. 1 basic position of the booster piston 16 is biased against a se in the Gehäu 2 inserted spindle 48, whose left in FIG. 1, annular end face 50 is conical inclined inwardly. The cone angle of the ring end face 50 is selected so that the translator piston 16 with its radially inner end face section rests on the ring end face 50 of the spindle 48 . Through this radially inner; spherically curved contact surface can be compensated for manufacturing tolerances, so that torque-free support of the translator piston 16 is ensured. This configuration of the translator piston 16 can significantly reduce the manufacturing outlay. By selecting the screw-in depth of the spindle 48 , the basic position of the booster piston 16 and thus the minimum preload of the compression spring 14 can be changed, so that a certain time correction of the switching delay can be carried out. The Axialpo position of the spindle 48 is determined by a lock nut 52 . The end position of the booster piston 16 is determined by an oblique shoulder 54 of the valve bore 6 ; onto which the outer circumferential area of the base body 38 of the booster piston 16 runs with an axial displacement to the left ( FIG. 1).

FIG. 2 shows an enlarged representation of the area in which the flow control valve 46 is accommodated. This flow control valve 46 has a pressure compensator piston 54 which is biased against a stop shoulder 58 of the axial bore 32 via a control spring 56 supported on the locking screw 42 . The pressure compensator piston 54 , which is designed as a hollow piston, has a measuring orifice 60 in its base, via which the axial bore 32 is connected to the interior of the pressure compensator piston 54 .

The peripheral edge of the rear ring face of the pressure compensator piston 54 lying on the right in FIG. 2 forms a control edge 62 , via which a radial bore of the booster piston 16 can be controlled. This radial bore 64 opens into a translator space 66 which is delimited by the annular end face 50 on the one hand and the end face 40 on the other hand. In the area of the radial bore 64 , this pressure transfer chamber 66 is enlarged by an incision in the axial direction, so that the entry of control oil through the radial bore 64 is facilitated. The controllable via the control edge 62 mouth cross section of the Ra dial bore 64 thus forms the control orifice of the flow control valve 46 . In its control position, the pressure compensator piston 54 throttles this control orifice ( 64 , 62 ) in such a way that the pressure loss across the measuring orifice 60 remains constant, so that the volume flow through the flow control valve 46 remains constant regardless of pressure and temperature.

The function of the directly controlled pressure limiting valve 1 according to the invention will first be explained with reference to FIG. 3, which represents the time course of a pressure rise at the input port P. In the event of a pressure increase at the axial inlet connection P, the valve spool 10 initially remains biased into its closed position by the compression spring 14 . The spring tension of the compression spring 14 is minimal since the booster piston 16 is in its contact position on the spindle 48 (see FIG. 1). Upon reaching a minimum pressure pmin lifts the valve spool 10 from the valve seat 12, so that pressure medium can flow from P to T. At the same time, control oil flows through the nozzle bore 26 , the interior 28 , the damping nozzle 34 , the axial bore 32 , the measuring orifice 60 and the radial bore 64 into the booster chamber 50 . This control oil flow is kept constant by the flow control valve 36 regardless of viscosity. The control pressure tapped at a distance from the outlet port P acts in the spring chamber 18 . As a result of the pressure force resultant, the booster piston 16 is moved to the left in the illustration according to FIG. 1, so that the pretension of the compression spring 14 increases and the control slide 10 is acted upon in the closing direction. By supplying a constant control oil volume flow into the booster chamber 50 , the booster piston 10 is continuously acted upon in the closing direction; so that the pressure at the inlet port P increases linearly. The maximum Fe dervorsension is reached when the booster piston 16 runs onto the inclined shoulder 51 of the housing 2 , so that the maximum pressure to be limited p _{max is reached} at the input port P when the booster piston 16 has moved into its end position. With a further increase in quantity, the valve slide 10 lifts off its valve seat and is brought into its control position in which the pressure at the input port P is kept constant.

When the pressure at the inlet port P or at Druckan rose at the outlet port T, the booster piston 16 is acted upon by the control oil pressure in the spring chamber 18 to the right, so that the control oil from the booster chamber 50 through the fully open radial bore 64 , the pressure valve abutting the shoulder 58 genkolben 54 , the orifice 60 , the axial bore 32 , the damping throttle 34 and the nozzle bore 26 flows to the input port P down. The valve spool 10 is then moved back into its closed position by the pressure in the spring chamber 18 - the pressure limiting valve is ready for the next cycle. The return movement of the booster piston 16 is extremely quick compared to the opening movement.

Fig. 1 shows a circuit diagram in which two are connected to a consumer, for example a hydraulic motor leading working lines A, B via two oppositely-connected pilot-operated pressure relief valves 1, 1 'with each other.

The pressure relief valve 1 described above is characterized by circuit symbols, the valve slide 8 acting in the opening direction on the one hand in the opening direction by the pressure at the input port P and on the other hand by the force of the compression spring 14 and the control pressure at the port Y in the closing direction, which determines the pretension of the compression spring 14 16 and the flow control valve 46 determining the control oil flow to the booster piston 16 are shown. The orifice plate 60 and the pressure balance piston 54 struck in the opening direction by the control spring 56 and the pressure downstream of the orifice plate 60 and in the closing direction by the pressure upstream of the orifice plate 60 are also symbolically represented.

In the relative arrangement shown in FIG. 4, the pressure line valve 1 assigned to the work line B is connected with its input port P to the work line B and with the output port T to the work line A. The pressure relief valve 1 'is connected with the input connection P to the working line A and with the output connection T to the working line B.

If the pressure in the working line A is exceeded, the pressure relief valve 1 'is opened, so that pressure medium flows into the working line B. This high pressure in the working line A is also on the pressure relief valve 1 via the outlet port T, so that the intensifier piston 16 is very quickly moved back to its starting position by the high pressure at the control port Y, so that in the event of a pressure reversal, that is, a pressure increase in the working line B, the switching delay shown is carried out while the pressure relief valve 1 'is moved back into its starting position by the rapid relief of the translator space 50 and is thus in a position so that a subsequent pressure build-up in the working line A is damped , That is, he inventive pressure relief valve 1 allows to dampen pressure peaks even in the fast len reversing operation, so that damage to the system is prevented.

As already mentioned at the outset, the applicant reserves the right to the three essential aspects of the present invention: the tapping of the spring chamber 14 which is removed from the output port T. Control oil, the provision of a flow control valve 46 , which is integrated in the booster piston 38 or the spherical design of the rear end face of the booster piston to judge independent claims or to make them the subject of divisional applications.

In principle, the design features described above can also be implemented with a valve in a seat construction. Another major advantage of the solution described above is that the structure is relatively insensitive to vibration, since the valve spool 10 is designed with a surface difference, which is defined by the diameter of the valve seat 12 and the guide diameter of the valve bore 6 - this annular surface also acts in the closing direction ,

The linear pressure increase shown in Fig. 3 prevents a drop in volume when other consumers are operated in parallel to the directly connected consumer. The construction according to the invention is characterized by a very high power density (200 l / min are possible with a seat with a diameter of 9 mm).

The pressure relief valve according to the invention can be used particularly well for slewing gear and Turas drives, and in principle closed and open hydraulic circuits are possible. The characteristic curve can be changed by varying the diameter of the measuring orifice 60 or by changing the spring rate of the compression spring 14 . Preliminary tests showed that the diameter of the radial bore 64 controlled by the pressure compensating piston 54 must be chosen to be relatively small in order to set the desired time delay. It turned out to be optimal if only a single, small radial bore was provided.

Disclosed is a pressure relief valve in which a valve rail is biased into a closed position via a compression spring. The Compression spring is supported on a booster piston, the backwards control room is pressurized. The pressure in which the compression spring receiving spring space is in a region of Tapped output port in which this control pressure is not is influenced by flow forces. The pressure medium supply in the The translator piston chamber can be viscous by means of a flow control valve be kept constant independently. Fault tolerances at ferti the translation piston can be achieved by spherical design its rear contact surface on a spindle for adjusting the Compensate the axial end position of the booster piston.  

LIST OF REFERENCE NUMBERS

1

Pressure relief valve

2

casing

4

manifold

6

valve bore

8th

valve slide

10

seat cone

12

valve seat

14

compression spring

16

Booster piston

18

spring chamber

20

Jacket bore star

22

annulus

23

location hole

24

bore star

26

nozzle bore

28

inner space

30

guide projection

32

axial bore

34

damping

36

radial shoulder

38

body

40

face

42

end

44

Screw

46

Flow control valve

48

spindle

50

Booster chamber

51

angular contact

52

mother

54

Pressure regulator piston

56

control spring

58

stop shoulder

60

orifice

62

control edge

64

radial bore

66

Booster chamber

Claims (9)

1. Directly controlled pressure relief valve with a valve spool ( 8 ) guided in a housing ( 2 ), via which an input connection (P) can be connected to an output exclusion (T) and which acts against the force = a compression spring accommodated in a spring chamber ( 18 ) ( 14 ) is axially displaceable, which is supported on a booster piston ( 16 ), the translator chamber ( 50 ) which is effective in the direction of increasing the spring preload can be acted upon by the pressure at the inlet port (P), characterized in that the spring chamber ( 18 ) is above a control connection (Y) with which pressure can be applied in the area of the output connection (T), the control pressure being tapped at a distance from the output connection (T). 2. Pressure relief valve according to claim 1, wherein the control connection (Y) opens at an axial distance from the outlet connection (T) in a receiving bore ( 23 ) for the pressure relief valve ( 1 ). 3. Pressure relief valve according to claim 1 or 2, wherein a flow control valve ( 46 ) is arranged in the pressure medium flow path between the input port (P) and the booster chamber ( 50 ). 4. Pressure relief valve according to claim 2, wherein a pressure balance piston ( 54 ) of the flow control valve ( 46 ) in an axial bore ( 32 ) of the booster piston ( 16 ) is mounted and by means of axial displacement of the pressure compensator piston ( 54 ) a nozzle in the jacket of the booster piston ( 16 ) can be controlled and a measuring orifice ( 60 ) is formed in the bottom of the pressure compensator piston ( 54 ). 5. Pressure relief valve according to claim 3, wherein the nozzle is formed by a single, small radial bore ( 64 ). 6. Pressure relief valve according to one of the preceding claims, wherein the volume of the spring chamber ( 18 ) is variable. 7. Pressure relief valve according to one of the preceding claims, wherein the booster piston ( 16 ) is supported abge on a spindle ( 48 ) and an end face portion of the booster piston ( 16 ) is spherically curved and the corresponding annular end face ( 50 ) of the spindle ( 48 ) is formed in this way is that the system takes place in the radially inner region of the end face ( 40 ) of the pressure booster piston ( 16 ). 8. Pressure relief valve according to one of the preceding claims, wherein the booster piston ( 16 ) has a guide projection ( 30 ) which is immersed in an interior ( 28 ) of the booster piston ( 10 ), wherein in the end face of the guide projection ( 30 ) has a damping nozzle ( 34 ) is formed. 9. Pressure relief valve according to one of the preceding claims, wherein the valve slide ( 10 ) is biased against a valve seat ( 12 ).