EP0883753A1

EP0883753A1 - Load-holding brake valve

Info

Publication number: EP0883753A1
Application number: EP97907055A
Authority: EP
Inventors: Hubert Häussler; Ivan Hristov; Hans Staiger; Josef ZÜRCHER
Original assignee: Beringer-Hydraulik AG; Beringer Hydraulik AG
Current assignee: Bucher Hydraulics AG
Priority date: 1996-02-28
Filing date: 1997-02-28
Publication date: 1998-12-16
Anticipated expiration: 2017-02-28
Also published as: JP3617841B2; DE59707059D1; US6098647A; JP2000505532A; KR19990087371A; WO1997032136A1; EP0883753B1

Abstract

The invention relates to a hydraulically controlled valve for double-acting consumers, which discharges during the lowering mode in a throttled manner a load volume flow from connection B to connection A. To this end, the valve has a valve seat (5) which comes into contact with a sealing surface (4) of a main piston (3) in which a pilot piston (8) is guided concentrically which comes into contact with a sealing surface (7) of a pilot valve seat (7) in the main piston. Said main piston (3) and said pilot piston (8) are held by load pressure and/or by resilient force on the valve seat (5) or the pilot valve seat (6). The pilot piston (8) has on the pressure-relieved side a piston shank (9) extending in a seat bore and having a continuously decreasing throttling action, in particular throttle grooves (10) which form a throttling point. When hydraulically actuated the pilot piston (8) can be axially displaced by a control piston (20) with the result that the pilot valve (5, 6) opens and there is a load volume flow via the throttling point from a pilot chamber (15) accommodating the springs. The resultant decrease in load pressure causes axially displacement of the main piston (3) and opening of the valve seat. An additional damping valve arrangement reduces load oscillations which occur. A pressure-limiting-valve independent of the return pressure is incorporated in the load-holding brake-valve housing.

Description

DESCRIPTION

Load holding brake valve

The invention relates to a hydraulically controllable load-holding brake valve for double-acting consumers according to the preamble of claim 1.

This valve is known from CH-PS 54 30 28. The load-holding brake valve has a ball seat valve as a pilot valve, which requires very sensitive control by the pilot piston in order to enable a steady increase in flow from the beginning and an abrupt change Avoid opening behavior.

It is therefore the object of the invention to provide a load holding brake valve of the above. Execute genus so that a very sensitive pressure relief of the main piston takes place without a sudden increase in volume flow. In order to achieve a progressive opening behavior of the valve, the load volume flow must be reduced very evenly and steadily.

A progressive opening behavior is understood here to mean that the opening is essentially proportional to the control pressure, but that in any case there is a defined, jolt-free and shock-free dependence between the opening cross section and the control pressure (ie: the 1st and 2nd derivation of the function: "opening of the valve above the control pressure "are finite and steady for every change in the control pressure).

This object is achieved by a load holding brake valve with the features according to claim 1. Here, the pilot piston has a piston shaft (pilot plunger) connected to its sealing surface, which is guided in the seat bore with little play. The piston skirt has a sequence of several cross-sectional areas along its length. The maximum cross section follows and corresponds to the seat cross section. It essentially has very little play compared to the seat bore (pilot channel). The area with the maximum cross section is very short and can go to zero.

This is followed by the throttle area. Through it, the valve tappet forms a throttle point in the seat bore (pilot channel), the throttling effect of which becomes smaller - and preferably progressively - smaller when the tappet is displaced and / or emerges from the pilot channel. At one point in the throttle area, preferably shortly after the pilot piston is lifted from its seat, the throttling effect is so great that it is substantially greater than the throttling effect of the compensating throttle. From there, the throttle effect with respect to the pilot control channel decreases with increasing control path of the pilot piston and displacement of the pilot control plunger in such a way that the function of the throttle effect is constant over the displacement length of the pilot control plunger. The throttle effect preferably decreases slightly at first and then decreases more and more with increasing displacement; The length and cross-section of the plunger are so matched that the plunger forms a very small throttle gap with the pilot channel of the main piston at the start of opening of the pilot piston, the throttle effect of which is substantially greater than the throttling effect of the compensating throttle and then until it is reached of the minimum cross-section forms a steadily increasing throttle gap, the throttle effect of which decreases with the length of movement of the pilot tappet and control piston and preferably more and more so that the closing forces acting on the main piston decrease. This is due to the special design of the throttle area with regard to its length and reached its cross section. In particular, two cross-sectional designs are possible:

In the first embodiment, the cross-section starts from the maximum cross-section and then decreases continuously over its length to the minimum cross-section. Here, the decreasing throttle effect is achieved in that the throttle gap of the pilot tappet increases steadily with respect to the pilot channel wall, starting from the throttle spatula of the maximum cross section.

In the second embodiment, the cross section also starts from the maximum cross section; it then decreases over a first partial length to a cross-section that is larger than the minimum cross-section; This cross-section remains the same over a further partial length. Here, the decreasing throttle effect is achieved in that the length of the throttle gap of the throttle valve plunger, which plunges into the pilot channel, steadily decreases with the displacement of the pilot piston. Combinations of both versions are also conceivable.

The constant reduction in the cross-section of the tappet in the throttle area, ie: the decrease in the throttle effect of the tappet relative to the pilot channel can be achieved, for example, in that the tappet in the area with a reduced cross-section is a rotating body, the diameter of which is slightly conical over its length or - preferably • progressive, ie: parabolically or hyperbolically reduced. It would also be possible to design the area with a decreasing cross section to be cylindrical with the diameter of the area with the maximum cross section, but in the area of the reducing cross section to provide a chamfer or flattening or axial throttles of variable depth and width, which differ from the area with the maximum Cross section going out and to the circumference of the area with the smallest cross-section. In order to achieve a progressive reduction in cross-section, the depth can additionally or alternatively also increase the width of these chamfers or throttle grooves parabolically or hyperbolically.

The length of the throttle region is preferably matched to the change in the closing forces acting on the main piston. These closing forces are made up of the hydraulic forces and the spring forces acting on the main piston.

This adjustment takes place in such a way that the tappet forms a very small throttle gap with the pilot channel of the main piston at the start of opening of the pilot piston, the throttle effect of which is greater than the throttle effect of the compensating throttle and then forms a continuously increasing throttle gap until the minimum cross section is reached. the throttling effect decreases with the movement length of the pilot plunger and control piston more than the closing forces acting on the control piston increase. In the embodiment of the valve with a spring acting directly on the main piston, the length of the throttle area is inversely proportional to the spring stiffness of the springs acting on the main piston in the closing direction. This configuration ensures that after the pilot valve opens, the pressure reduction in the pilot chamber begins slowly and depends on the length of the tappet and the length of movement of the control piston or pilot tappet on the one hand and the hydraulically effective design of the main piston and the spring stiffness the springs that push the main piston in the closing direction, on the other hand. The greater the spring stiffness of these springs, the shorter the plunger and its throttle block. In other words: the more the spring force on the main piston increases when opening, the stronger must be decrease the throttling effect in the pilot channel when the pilot tappet moves.

This results in a stable balance days of the main piston at each opening pressure. A sudden, jerky and abrupt opening or movement of the load-holding brake valve piston is thus prevented. The opening characteristics and the adjustment of the ram length on the one hand and the springs on the other hand also prevent the excitation of vibrations on the movable loads.

The pressure reduction in the pilot control chamber takes place continuously and in a defined dependence on the control pressure and the movement of the control piston caused thereby. It is avoided that the main piston leads the pilot piston or executes uncontrolled and uncontrollable movements.

A precisely working hydraulic follow-up system is created. As soon as the load pressure behind the main piston is reduced, the main piston automatically follows the pilot piston, since the load pressure, which acts on the resulting annular surface of the main piston, pushes the main piston out of its valve seat. Since the main piston follows the pilot piston displaced by the pilot piston, the opening cross section in the pilot valve seat will in turn narrow, so that a back pressure can build up in the pilot chamber of the main piston. An equilibrium position is thus established between the control piston and the pilot piston on the one hand and the main piston on the other. The advantage of this principle is that the flow force acting in the closing direction is smaller than the hydraulic opening force under all conditions due to the hydraulic force amplification achieved. This prevents pressure vibrations in the consumer port B from triggering undesired movements of the main piston. This can do that A boom on the hydraulic excavator or crane can be prevented from rocking.

The maximum cross section of the pilot tappet is designed relative to the cross section of the pilot channel so that the throttle cross section after opening the pilot valve seat is initially only slightly smaller than the throttle cross section of the compensating throttle, so that when the main piston is lifted from the seat, a slow pressure reduction in the pilot chamber can take place. By designing the area with the maximum cross-section and the adjoining area with the reduced cross-section, various opening characteristics can be realized, in particular also an opening behavior that is linear to the control pressure and linear to the movement of the control piston.

Overall, the load holding brake valve according to the invention integrates the functions for load holding, lowering, load lifting and load securing in a valve housing with a very compact design. It is achieved that with a correspondingly designed course of the throttle cross section, a hydraulic force constantly acts on the pilot piston, so that this or the plunger cannot lift off the pilot piston, despite the pressure drop on the spring side of the pilot piston.

In the load lifting state, the spring-loaded main piston also acts as a check valve. The small opening pressure of the check valve is made possible by a large seat area. The large area ratio between the effective diameter of the valve seat and the effective diameter of the pilot valve seat means that the pilot piston does not open. The invention allows a fine gradation of the various throttle points which are formed or are present in the valve. In the embodiment of the valve according to claim 3, the control pressure is largely independent of the load pressure to be controlled, so that a wide control range and sensitive control is made possible even with a low control pressure.

The design according to claim 4 ensures that the opening of the valve seat, which leads to an imprecisely definable throttling due to the size of the valve seat diameter, has no negative influence on the opening behavior of the valve.

The design according to claim 5 ensures that the control with low control pressure already leads to a reaction of the valve and a movement of the load.

According to the invention, it is possible to guide the opening and closing movement of the main piston hydraulically alone, since there are always definable hydraulic states on both sides of the main piston due to the fine-tuning of the throttle points.

The embodiment according to claim 3 serves to increase the closing effect, wherein two springs connected in parallel can preferably be provided for security. Of these springs, one can only act on the main spool, the other, however, on the pilot spool and thus indirectly also on the main spool. Because of the sensitive pilot control of the main piston, it is advantageously also possible to load only the pilot piston with a spring in the closing direction, which at the same time also acts on the main piston in the closing direction. With regard to the formation of the pre-throttle bore in the guide shaft, there is extensive design freedom, depending on the desired behavior. The cross-section of the pre-throttle bore can be greater than / equal to or smaller than the flow cross-section of the compensating throttle between the pilot control chamber and the annular chamber.

The embodiment according to claim 6 is advantageous if the control pressure is specified independently of the inlet pressure. The graduation of the pilot piston and the provision of a pilot throttle have the effect that a higher closing pressure acts on the pilot piston. The flow from B to A can thus be reduced with increasing load pressure, particularly in open systems with external control. The reaction of the valve is damped, so that irregularities or vibrations of the load or discontinuities in the control, in particular, cannot act as vibrations of the valve.

Because the pilot piston and the pilot piston are guided independently of one another, misalignments between the pilot piston and the pilot piston have no effect.

In the load holding brake valve according to claim 7, a pressure relief valve for securing the load pressure is integrated in the valve housing. In addition, the highest load pressure can be set in a simple manner.

It is expedient to take measures to reduce the spring load on the pressure limiting piston. This goal is achieved in that the pressure limiting piston has only a small effective area which is acted upon by the load pressure. As a result, the required spring force is greatly reduced and the installation space is reduced. Such an embodiment results from claim 8.

For safety reasons, two pressure springs act on the pressure-limiting piston, which are connected in parallel, but which are arranged in accordance with claim 9 to save space.

In the design of the load holding brake valve according to one of claims 7 to 9, the opening pressure of the pressure relief valve is also independent of the return pressure. This principle of the pressure relief valve ensures that there is no summation of the set pressures in a pressure relief valve connected downstream in the directional control valve.

As already stated, the load holding and braking valve according to this invention has a pilot piston with which the pilot piston is actuated hydraulically and mechanically. Such hydraulic-mechanical actuation of valves occurs frequently in hydraulics, for. B. in the hydraulic actuation of control valves. This hydraulic control has the disadvantage that the opening of the control valve leads to a - more or less rapid - increase in the control oil quantity. It therefore depends on the attentiveness and skill of the operator that the control valve when a desired control oil quantity is reached, ie. H. when a certain position of the hydraulically-mechanically controlled valve is reached, to keep the control valve in this position.

Another object of this invention is to bring this, like any other hydraulic-mechanically controlled valve, into a predetermined end position. Such an embodiment results from claim 10. Such metering valves can be based on a hydraulic principle of action, for. B. by metering predetermined Ölmcngen, which are then fed as control oil to the control piston.

However, this would make it necessary to adapt the amount of oil to be metered to the desired path of the control piston. This adaptation occurs automatically when the valve is designed according to claim 11.

Here, the locking element can be mechanically contacted with the pilot piston in any way, so that the supply of a further pilot oil flow is interrupted when a predetermined position of the pilot piston is reached.

A hydraulically and mechanically advantageous integration of the metering valve into the opening device results from claim 12. In this embodiment, there is an easily accessible and easy-to-use possibility of adjusting the metering amount from claim 13. Furthermore, an emergency actuation of the opening piston is desired and achieved by the Embodiment according to claim 14, 15 and 16. In this way, mechanical actuation of the control piston upon failure of the control pressure and, on the other hand, complete deactivation of the controllability of the control piston are effected. Both functions can be offered for security reasons.

Claim 17 provides a simple way of relieving the pressure of the closing element.

Claim 18 describes further execution options. These designs thus on the one hand enable the control to be limited in terms of the stroke, and on the other hand the possibility of a mechanical response, in particular as an emergency function.

The stroke limitation very often does not represent the desired end position of the control piston and the valve actuated thereby, but rather only a position from which the end position is to be approached. This is particularly the case when lowering loads. There, the main route is to be traversed at high speed, while the end position is slow, i. H. is started in slow speed.

Claim 19 forms the valve according to claims 10 to 18 accordingly. With this configuration, the fast mode can be switched on very suddenly by operating the valve when the metering valve is open. After the end stroke of the metering valve has been reached, on the other hand, operation in creep speed takes place via damping nozzles which allow the speed of the creep speed to be set. The now damped operation of the control piston allows the desired end position to be reached by fine control. The ratio of the overdrive range to the final gear (fine control range) can be adjusted externally by adjusting the adjustment spindle for the metering valve. The quick reaction of the plunger remains possible despite the strong hydraulic damping. In order to enable a quick reaction even outside the functional range of the metering valve, a preload valve is provided which opens the connection of the control pressure channel to the control chamber and control piston when a certain control pressure is exceeded (claim 20).

The embodiment according to claim 21 serves the purpose of damping pressure fluctuations in the control pressure both in the high-speed mode and in the fine control mode. Further advantages and exemplary embodiments of the load holding brake valve according to the invention will now be explained in detail with reference to the drawings.

Show it:

Fig. 1 hydraulic circuit diagram for controlling a consumer in

Meaning an adaptation of a discharge flow to an inlet flow,

2 shows a longitudinal section of a variant of a load holding brake valve,

3a / b shows a longitudinal section through a pilot valve,

4 shows a longitudinal section of a variant of a load holding brake valve,

5 hydraulic sound plan corresponding to FIG. 1 with hydromechanical limitation of the control current,

6 detail of the metering valve of FIG. 5,

Fig. 7 shows the detail of Fig. 6, but in the state of the extended

Stroke of the control piston, d. h .: with hydraulic stroke limitation

Fig. 8 detail of FIG. 5, but with mechanical control of the

Control piston,

FIG. 9 hydraulic circuit diagram corresponding to FIG. 1 with a hydraulic damping of the discharge flow, with pressure relief valve, damping bypass and relief bypass. 10 embodiment of FIG. 9th

In Fig. 1, the hydraulic circuit diagram of a control of a consumer is shown in the sense of an adaptation of a drain flow to an inlet flow by means of a load holding brake valve. The consumer 26 is connected to the inlet line 28 and the lower line 25. The lowering line 25 is connected to the connection B of the load holding brake valve 1A. A return line 27 leads from the load-holding brake valve 1A from the connection A to the directional control valve 31. The supply line 28 also ends at the directional control valve 31. The directional control valve 31 is designed as a 4-3-way valve. In addition to the inlet line 28 and the return line 27, the connection of a pump 32 and the connection of a line to the tank 33 are provided. The load holding brake valve 1A is connected to the inlet line 28 via a control line 29. Furthermore, the pressure relief valve 30 is connected between the lower line 25 and the return line 27. In the switch position shown, the inlet line 28 and the return line 27 are connected to the tank 33. The consumer 26 thus remains in the current position. The connection between port B and port A of the load holding brake valve 1A is blocked.

If the slide directional control valve 31 is moved to the right, the inlet line 28 is connected to the pump 32. The consumer 26 is now in the lowering mode. Therefore, the return line 27 is connected to the tank 33. However, the connection between the connection B and the connection A in the load holding brake valve 1A remains closed until the pressure has built up in the inlet and a control pressure sufficient via the control line 29 is present at the load holding brake valve 1A. The load holding brake valve 1A is then shifted to the right against the spring. Port B and port A in the load holding brake valve 1A are now connected to each other via a variable throttle. The volume flow thus flows out of the Senklcitung 25 to the return line 27 and in the tank 33. The load holding brake valve 1A remains in this position as long as the control pressure is constant. Every change in the inlet pressure thus has a direct effect on the opening cross-section of the load holding brake valve.

If the slide directional valve 31 is shifted to the left, the pump 32 is connected to the return line 27. The feed line 28 is connected to the tank 33 so that there is no pressure on the control side of the load holding brake valve 1A and the load holding brake valve 1A remains in the position shown. In this position, the consumer 26 is in the lifting mode. The pump volume flow reaches the port B via the return line 27 and the check valve in the load holding brake valve 1A. From there, the oil flows through the lower line 25 to the consumer 26.

The pressure relief valve 30 serves to secure the load pressure in the lowering mode or when the consumer is at a standstill and is arranged between the lowering line 25 and the return line 27. An additional pressure protection is usually arranged on the directional valve (not shown here).

2 shows a longitudinal section of a load holding brake valve without an integrated pressure relief valve. The load holding brake valve has a housing 1 with a cylindrical control chamber 2. The control chamber consists of, preferably aligned, chamber sections in this order: pilot chamber 15;

Annulus 70, which is connected (via connection B) to the lowering line 25 of the consumer 26, return space 73, which is connected (via connection A) to the return line 27 to the tank; Control chamber 21, which is connected to a control channel X;

The cylindrical control chamber 2 is closed at the end by means of a control chamber plug 13. The connection bores A and B open into the control chamber 2 perpendicular to the longitudinal axis of the control chamber 2. Between the connection bores A and B, the control chamber 2 has a valve seat 5. The valve seat 5 is fixed in place on the valve housing 1 and separates the annular space 70 from the return space 73. Between the one end of the control chamber 2 with the plug chamber plug 13 and the valve seat 5, a main piston 3 is movably guided. The main piston 3 has a thinner collar with a conical sealing surface 4 which interacts with the valve seat 5. The main piston 3 has an end collar 42 on the side facing away from the valve seat 5 and facing the connection bore B. The end collar 42 hal has a larger diameter than the above-mentioned collar and is sealingly guided in the control chamber 2, so that the main piston 3 is axially movable. Due to the design as a stepped piston, the main piston 3 forms the annular space 70, which is connected to the lowering line 25 via connection B. The annular space 70 is connected to the return space, the connection B and the tank 33 by lifting the main piston 3 from the valve seat 5.

The area of the control chamber 2 between the thick end collar 42 of the main piston 3 and the plug chamber plug 13 is referred to as pilot chamber 15. This Vorstcuerraum 15 serves to receive a spring 12A (not shown), which is clamped between the control chamber plug 13 and the main piston 3. Shown is a spring 12, to be described in more detail later, which has the same function in that the main piston 3 is pressed onto the valve seat 5 by spring force, but also by the hydraulic forces acting on it. The annular space 70 is connected to the pilot control space 15 via the throttle 14. The throttle 14 can - as shown - be arranged axially parallel in the thicker piston collar, but also in the valve housing. The main piston 3 is penetrated concentrically by a pilot channel 34, which connects the pilot chamber 15 to the return chamber 73. For this purpose, the main piston 3 has a step bore 71 which is arranged concentrically with respect to the pilot chamber 15. The step 72 designated as pilot channel 34 with a smaller diameter starts from the bottom 72 of the first step. The pilot valve seat 6 is formed on the base 72 between the stage 71 and the pilot channel 34.

A pilot piston 8 is guided with its piston shaft, the pilot plunger 9, in the pilot channel 34 with play. Pilot piston 8 and pilot plunger 9 are made from one or two pieces. The pilot plunger 9 has a smaller diameter than the pilot piston 8 protruding from the pilot duct 34. The pilot piston 8 has at its end, which is connected to the pilot plunger 9, a sealing surface 7 which acts under the force of the pilot spring 12 (closing spring) the pilot valve seat 6. The smaller truncated cone surface corresponds essentially to the cross section of the pilot channel 34 and the adjoining area of the pilot valve tappet 9.

The pilot tappet 9 has a plurality of diameter areas or cross-sectional areas over its length.

A small groove adjoins the conical seat as an undercut. The groove runs in the circumferential direction and for essentially technical reasons. A very short area of large cross-section (= cross-sectional area) of the pilot piston 8 adjoins the groove. This area is cylindrical and, with little play, has a diameter which corresponds to the diameter of the pilot channel 34 and the smaller sealing surface 7. Its length can go to zero so that it only represents the beginning of the following range.

The very short area of large cross-section is followed by an area with a decreasing throttling effect. The throttling effect which decreases with the tappet movement is achieved in that the cross section of this area - starting from the maximum cross section - is continuously reduced at least over a partial length and / or that the part of this partial length which is immersed in the pilot channel changes when the Input ram shortened. Another part of this area can have a constant cross-section, which is, however, larger than the cross-section of the following area with the smallest cross-section. The decreasing throttling effect results from the fact that with the tappet movement the area with a decreasing cross-section emerges from the pilot channel into the pilot chamber. With a further tappet movement, the part length of the pilot tappet 9 immersed in the pilot channel 34 changes and has a constant cross section. The change in the throttling effect in this area of the pilot tappet 9 thus takes place by changing the throttle cross section and / or the throttle length, which plunges into the pilot channel 34 and is guided therein. This means that the area with a decreasing cross section or decreasing throttling effect may not be longer than the pilot channel 34. The length depends in particular on the desired opening behavior in relation to the control pressure.

However, the area can be made cylindrical with the diameter of the preceding area and chamfers or on the cylinder jacket Make grooves that start from the largest cross section and end on the smaller, constant cross section. Flow and production engineering favorable designs of the area with decreasing cross-section are described with reference to Figures 3a and 3b.

The end of the pilot plunger 9 (piston skirt) has a smallest cross section, which essentially corresponds to the smallest cross section of the area with a decreasing cross section. This end area is only partially within the pilot channel 34. It extends beyond the length of the pilot channel 34 and protrudes with its end into the return chamber 73 of the control chamber 2.

In particular, it is shown that the pilot tappet 9 has a circumferential Frcisiich groove 35 following the sealing surface 7. This is followed by a cylindrical area, the diameter of which corresponds to the diameter of the pilot channel with play (area with the largest cross-section, area with the maximum cross-section).

The "area 143 with decreasing throttling effect" begins at a short distance from the undercut. The entire area 143 can have a decreasing cross section and can be designed as a rotating body with a straight or preferably parabolic or hyperbolic surface line.

Special attention must be paid to the transition between the area with the maximum cross section and the area with a decreasing cross section. This transition must be steady so that when the pilot tappet 9 moves in this area there is no impact or jerk in the load movement.

In Fig. 3a, the decreasing throttling effect of the area 143 on the first part length 144 is shown by a decreasing cross section of the tappet reached. The ram is weakly conical over this part length, ie: formed as a truncated ball. The large conical surface corresponds to the cross section of the preceding area with the largest cross section. The small conical surface corresponds to the cross section of the subsequent partial length 145, which has a constant cross section. This partial length 145 produces only a slight throttling effect in the pilot channel, which decreases steadily with the appearance of this partial length from the pilot channel. This partial length is therefore only of minor importance for the function of the valve. Their length can therefore go to zero. What is important is the design and length of the preceding area with a decreasing cross section. It must be emphasized in relation to the drawing that the transition between the area with the maximum cross section and the area with decreasing cross section cannot be adequately represented there. In reality, no circumferential edge may be formed there, since a continuous, ie parabolic or hyperbolic, transition is desired. The parabolic or hyperbclical configuration of the surface line cannot be represented either. A linear mantle line is shown, which, however, cannot be considered preferred in the sense of this invention.

An area 146 with the smallest cross section adjoins the partial length 145 with a constant cross section. It should be emphasized that this smallest cross section is in any case smaller than the cross section of the preceding part length 145 with a constant cross section. In any case, the boundary between the two cross-sectional areas lies in the pilot channel when the pilot valve is closed. The area with the smallest cross-section protrudes from the pilot channel into the return space.

The embodiment of the pilot tappet 9 shown in FIG. 3b has in the

Area with decreasing throttling effect in the axial direction has a plurality of throttle grooves 10 which, with the wall of the pilot channel 34, the throttle place 36. In the region with a decreasing cross-section, the throttle grooves 10 have a depth that increases steadily and - preferably progressively - towards the free end of the pilot tappet 9 (partial length with a decreasing cross-section). Then they maintain the maximum depth reached (partial length with constant cross-section). Following the area with the throttle grooves 10 (area with decreasing throttling effect), the area with the smallest cross section also follows here. This area is again cylindrical. The diameter can essentially correspond to the diameter of the deepest groove base of the throttle grooves 10.

The throttle grooves 10 can be replaced by flats or notches, which are introduced axially or in a helical manner on the pilot tappet 9. Instead of or next to the depth, the width of the throttle grooves 10 can be changed. This is especially true in the initial area of the grooves, i. H. : in the area with decreasing throttling effect. The grooves start from the area with the maximum cross section with depth = 0 and width = 0. By increasing the width and depth of the grooves, a constant parabolic or hyperbolic or other course of the cross section of the ram can be achieved.

About the function:

When the pilot piston 8 is axially displaced to the right, the pilot tappet 9 opens by the sealing surface 7 lifting off the pilot valve seat 6. As long as the largest cross-sectional area is immersed in the pilot channel 34 (throttle point 36), the volume flow remains strongly throttled, this throttling effect, in comparison to the pilot throttle 14, determining the pressure drop in the pilot chamber and thus the opening behavior of the main piston. With increasing axial displacement of the pilot tappet 9, the area of the largest tappet cross section emerges from the pilot channel 34 and therefore continuously reduces its throttling effect. The throttling effect is now determined by the area with decreasing throttling, ie: first by the decreasing cross section of the pilot tappet 9 emerging from the pilot channel 34. Here the depth of the throttle grooves increases (Fig. 3b) and the diameter of the tappet decreases (Fig. 3a). The throttling effect becomes steadily less when this partial length (truncated cone or grooves) emerges from the pilot channel 34 into the pilot chamber. When the smallest conical cross-sectional area of the truncated cone or the greatest depth of the throttle grooves have reached the pilot valve seat 6, the decrease in the throttling effect continues nonetheless to a much lesser extent, since the length of the part immersed in the pilot channel decreases with a constant cross-section. The fact that the area with the smallest cross section remains immersed in the pilot channel 34 has no effect, since the throttling effect of this area is very small.

There is therefore a continuously slow pressure reduction. Here, the opening cross section of the pilot control seat 6, even immediately after opening, is larger than the throttle cross section in the pilot channel (throttle point 36).

A separator 17 (Fig. 4) separates the return space 73 from the axially aligned control bore 43. The control bore 43 is closed on the other end by the plug 22. A control piston 20 (guide collar) is sealingly guided in the control bore 43. The control piston 20 divides the bore 43 into the control chamber 21 and the spring chamber adjacent to the separating web 17. The plug 22 has a connection bore X through which the control chamber 21 is connected to the control line 29 (FIG. 1). The control piston 20 has an control shaft 16, 19, which consists of a thicker part 19 and a thinner part 16. The thinner part 16 of the control shaft penetrates the separating web 17 and is guided in the separating web in the guide bore 74 in a sealing manner (seal 18 or a sealing gap). The free end of the control shaft 19 with the end face 44 protrudes into the return chamber 73, the control shaft 16, 19 and the pilot plunger 9 of the pilot piston 8 lying on an axis line. The opening piston 20 is pressed into its starting position by a opening spring 24 designed as a compression spring, which is arranged in the spring chamber 43 and is supported on the separating web, when there is no control pressure in the opening chamber 21. The spring chamber 43 is relieved of pressure by means of the leak oil hole L. For safety reasons, the opening spring 24 is formed by one or more springs 46, 47 connected in parallel (see FIG. 4).

The thicker area 19 of the control shaft 19 forms an end face 48 compared to the thinner area 16. This end face serves as a stop face 48 for mechanically limiting the stroke of the control piston 20 by coming into contact with the separating web 17. For dimensioning, it should be said that the guide collar of the plunger 20 has an end face 45 which is acted upon by a control pressure, the effective area c of which is greater than 50: 1, preferably greater than 100: 1, relative to the effective area of the pilot control valve 6.

Furthermore, the ratio of the end face 45 on the guide collar to the end face 44 on the opening end 16 is greater than 30: 1, in particular greater than 60: 1.

With the advantageous embodiment described above, the control pressure remains largely independent of the load pressure. In particular, the control pressure also remains largely independent of the return pressure. The pressure relief of the spring chamber 43 in the region of the control spring 24 thus enables an exactly predetermined force curve acting on the control piston 20 and dependent on the build-up of the control pressure. For safety reasons, it is advantageous here if several springs act on the control piston 20. This ensures that, even if a spring breaks, the control piston 20 /. B. is still controllably displaced into its closed position in the event of a line break.

The course of the throttle cross-section on the pilot tappet is designed such that a displacement movement of the pilot piston 8 impressed by the control piston 20 in the opening direction is only possible with increasing hydraulic force on the control piston 20 for the lifting operation.

The ratio of the effective areas of the main piston 3 and pilot piston 8 is designed such that no relative movement between the main piston 3 and pilot piston 8 in the sense of an opening of the pilot valve seat 6 can be carried out.

About the function of the load holding brake valve: At standstill:

The load pressure of the consumer is present in the connection bore B and the annular space 70. The pilot control chamber 15 is connected to the annular chamber 70 via the throttle 14. The load pressure acts on the active surface of the thicker end collar 42 of the main piston 3. The main piston 3 with its sealing surface 4 is pressed hydraulically against the valve seat 5 by the spring 12. The pilot piston 8 is acted upon by the load pressure and the spring force of the spring 12; it is held with its sealing surface 7 on the pilot valve seat 6. The connection from B to A is thus blocked leak-free.

In lowering mode:

The directional control valve 31 (FIG. 1) connects the consumer 26 via the inlet 28 to the pump and via the return line 27 to the tank. The load holding brake valve is connected to the pump via the control line 29 and the connection bore X via the inlet 28. The pressure which can be changed by the directional control valve acts as a control pressure on the control piston 20. In accordance with the control pressure, the control piston 20 is displaced towards the separating web 17 against the control spring 24 until the spring force and control force are in equilibrium. The pilot shaft 16 abuts the free end of the pilot valve 9 of the pilot piston 8 with its end face 44 and displaces the pilot valve 9 - in absolute terms - by a distance that is proportional to the pilot pressure. The sealing surface 7 of the pilot piston 8 is lifted out of the pilot valve seat 6. This creates the connection between the return chamber 73 and the pilot chamber 15, the throttle effect of which depends on the design of the pilot plunger 9 and the length of the plunger path or control path or the level of the control pressure. At a low pilot pressure, ie: as long as the area of the pilot stroke 9 with the largest cross-section is inside the pilot channel 34, this connection is throttled very strongly. When opened further, however, the throttling effect becomes less than the throttling effect of the compensating throttle 14 in the main piston. This results in a slow pressure drop in the pilot chamber 15 and thus begins a slow movement of the main piston in the sense of the opening of the main valve seat 4 and the connection between the annular space 70 and the return space 73. The load therefore lowers very slowly. The movement of the main piston 3 in the sense of the opening of the main valve seat means relative to the pilot piston and tappet 9 move in the sense of closing the pilot valve 6/7, since the absolute position of the pilot tappet 9 is predetermined by the position of the pilot piston 20. Since the main piston 3 follows the pilot piston 8 in its movement, the throttle cross section on the throttle parts 36 in the pilot channel 34 will narrow again. As a result, a higher pressure builds up again in the pilot control chamber 15. This pressure build-up ensures that an equilibrium state is set between the pilot piston 8 and the main piston 3.

As soon as the control pressure is increased further, the area of the pilot tappet 9 with a decreasing cross section emerges further from the pilot channel 34 and the pilot valve seat 6. The pilot channel 34 is thus opened further, i. h .: the throttling effect of the pilot tappet 9 is further reduced. An increasing volume flow flows from the pilot chamber 15 past the pilot plunger 9, e.g. B. through the arranged in the pilot tappet 9 throttle grooves 10, in the return chamber 73. Here, the throttle cross section in the pilot channel 34, / .. B. through the throttle grooves 10, dimensioned such that with the movement of the pilot tappet 9 a smooth slow degradation the throttling effect and consequently a steady decrease in pressure in the pilot chamber 15 occurs. In this way, a progressive opening behavior of the pilot piston 8 is achieved, which is clearly defined by the size of the control pressure.

The length and throttle effect of the throttle area of the pilot tappet 9 are matched to the spring forces and hydraulic forces on the main piston 3. The main piston 3 immediately and evenly follows each movement of the control piston 20 and the pilot piston 8 and pilot piston 9.

The design of the main piston 3 in connection with the valve seat 5 also has the advantage that the flow acting in the closing direction always counteracts a hydraulic opening force that is greater than the flow forces in any position. The effects of possible pressure vibrations in port B on the main piston 3 are thus avoided.

Since the control piston 20 has a large effective area in relation to the pilot valve seat 6, the control pressure is essentially independent of the load pressure. The ratio between the effective area of the control piston 20 and the effective area of the pilot valve seat is greater than 50: 1, preferably greater than 100: 1. The control piston 20 has a ratio of its end faces 45 and 44, which is preferably greater than 30: 1. The pilot pressure is thus largely independent of the return pressure.

If the control pressure acting on the end face 45 in the Aufsteuerkam¬ mer 21 or - z. B. due to a line break - collapses, the control piston 20 is pushed back by means of the spring 24 and finally comes to a stop in its starting position. The pilot piston 8 follows it through spring 12 and closes the pilot valve seat 6/7. As a result, the pilot control valve is built up again in the pilot control chamber 15, which leads to the main piston following and closing the valve seat 4/5. The connection from connection hole B to connection hole A is closed so that the load of the consumer comes to a standstill.

In lifting mode:

Here, as shown in FIG. 1, the connection A is connected to the pump 32. The pump pressure picks up in the return chamber 73 on the valve seat 5, and lifts the main piston 3 against the spring force (spring 12 and possibly spring 12A) and opens the valve seat 5. The load is lifted. Due to the large difference between the effective area of the valve seat 5 and the The active surface of the pilot valve seat 6 will move the main piston 3 together with the pilot piston 8 in this check valve function. Due to the large area of the valve seat 4 on the main piston 3, only minor throttling losses occur at the valve seat.

It is pointed out that in the load holding brake valve according to the invention the compensating throttle 14 and the pre-throttle bore 41 can also be replaced by nozzles, so that a pressure-independent pressure reduction can take place.

A pressure limiting valve for load securing can be integrated into the load holding brake valve. This is shown and described with reference to FIG. 4.

The embodiment of FIG. 4 with respect to the control chamber 2 and control bore 43 as well as the valve function is identical to the valve Lasthalte¬ Fig. 2. Therefore, aul ^"is the description there Referring and only the differences noted.

In this exemplary embodiment, the preliminary piston 8 and the main piston 3 are advantageously clamped only with the spring 12, which is supported on the valve housing. The main piston 3 is moved axially essentially by hydraulic forces. In this embodiment, the pilot piston 8 has a guide shaft 37 which is guided in the stepped bore 71 of the main piston 3 in a sealing manner. A pre-chamber 40 to the pre-control chamber 15 is thus formed between the pilot valve seat 6 and the guide shaft 37 concentrically with the control piston 8. The pre-chamber 40 is connected to the pilot room 15 via a pre-throttle 41. The throttle cross section of the pre-throttle 41 can be made larger, equal to or smaller than the throttle cross section of the compensating throttle 14. This configuration of the .8th -

Pilot piston 8 has the advantage that the pressure reduction in the pilot chamber 15 takes place over two stages which have a fixed throttle cross section. In the open state in particular, the pre-throttle bore 41 has the effect that a higher closing force acts on the pilot piston 8 when the Lasidruck increases. The higher closing force means that the throttle cross section in the pilot channel (throttle position 36 in FIG. 3) is also reduced due to the axial displacement and thus the main piston 3 increasingly closes due to the follow-up control. This system is particularly advantageous in the case of an open circuit. In this case, a control pressure on the control piston 19 is predetermined, which is independent of the pump pressure and independent of the inlet pressure, eg. B. is also permanently adjustable.

Because of the large surface area of the control piston 20, two parallel biased springs 46 and 47 can be clamped between the guide collar 20 and the separating web 17 in the control bore (spring chamber) 43 as the opener spring. If one spring breaks, the other spring is able to move the plunger in its starting position. This is of particular importance in terms of security.

In the load hall brake valve shown in FIG. 4, a pressure limiting valve 30 is integrated in the valve housing 1. The pressure limiting valve 30 is designed as a check valve with permeability from the load side (annular space 70) to the tank scite (return space 73). The pressure-limiting piston 55 has only a very small area acting in the opening direction. This is achieved in that the pressure-limiting piston 55 has a shaft which penetrates the load chamber 53 and forms it as an annular space; that the load chamber 53 is delimited on the one hand by the piston 55 with a check valve seat 54 and on the other hand by an end collar 62 attached to the shaft, and that the check valve seat 54 has only a slightly larger hydraulic effective area than that on End collar 62 fastened to the shaft.

For the construction, a blind hole 50 is made on the end side of the valve housing 1 on the side of the control chamber. The blind bore 50 is connected to the annular space (load chamber) 70 by means of an overload bore 49 and to the return chamber 73 via the return bore 60. The plug 51 (bushing) is screwed into the blind hole 50. An inner bore 52 is made in the center of the stopper 51, which is open toward the blind hole and forms the check valve seat 54 with its end. The check valve seat 54 lies between the overload bore 49 and the return bore 60. The inner bore 52 is connected to the overload chamber 49 via the radial bores 53 and a recess 76 on the plug 51. The overload chamber 49 and the return chamber 60 are arranged between the bore 68 and the inner bore 52 of the valve housing of the pressure relief valve 30. The spring-loaded pressure-limiting piston 55 of the pressure-limiting valve 30 has a sealing surface 56, which bears against the check valve seat 54 under the biasing force of a compression spring 57, 66 and seals the radial bore 53 with respect to the return chamber 73. The pressure-limiting piston 55 has an end collar 62, 63 on both sides. The piston switch penetrates the radial bore 53 and has an end collar 62 at its end. This end collar 62 is sealingly guided in the inner bore 52 (seal 79), the end face 64 of which is somewhat smaller than the cross section of the check valve seat 54 of the piston. The end collar 63 attaches to the pressure limiting piston 55 and - with a tapered end part - is guided in the end wall and guide bore 77 with a seal 61 and projects into the bore 68. The inner bore 52, that is adjacent to the overload bore 49 and its end collar 62 are loaded with the pressure of the return chamber 73. A relief channel 81 is used for this purpose, which is designed as a longitudinal bore in the axis of the piston and which connects the return chamber 73 to the end space on the end collar 62 by means of a radial branch channel 80. The cross-section of this end space and of the end collar 62 is slightly smaller than the seat surface 54 of the check valve seat 54. The active surface, which is effective in the opening direction at load pressure, corresponds to this difference. The bore 68 is connected to the control bore 43 (spring chamber) and the leak oil bore L through the relief bore 69 for pressure relief. The thinner end collar 63 projecting into the bore 68 is of the same size with respect to its hydraulically effective cross-section (end face 65) as the above-mentioned active surface in the opening direction, ie: difference between the valve seat surface 54 and the cross section of the inner bore 52 with end collar 62.

The piston 55 of the pressure relief valve 30 is loaded with two pressure springs connected in parallel in the closing direction; one compression spring 57 is clamped in the return chamber against the piston shoulder 58 and the other compression spring 66 in the pressure-relieved end chamber against the piston skirt with end collar 63. To adjust the laser safety pressure, the plug 51 is screwed more or less deeply into the blind hole.

About the function of the pressure relief valve 30:

The load pressure is present in the inner bore 52 against the sealing surface 56 of the valve seat 56. As soon as the set load-securing pressure is reached without a volume syrup being removed via the main piston 3, the pressure-limiting piston 55 is axially displaced against the springs 57 and 66. The sealing surface 56 rises from the check valve seat 54 and the pressure relief valve 30 opens. The oil can now flow from the overload bore 49 through the open valve seat 54 to the return bore 60 stream. As a result, the annular space 70 and the return chamber 73 are connected to the main piston 3 bypassing the valve seat when the load pressure exceeds a preset limit value. The limit value (load-securing pressure) is specified by the two pressure springs 66 and 57 connected in parallel in series.

The design of the pressure limiting piston 55 and its pressure relief means that the opening pressure, which acts on the valve seat 54 in the inner bore 52, is independent of the return pressure and exclusively dependent on the load pressure. This exemplary embodiment of a pressure limiting valve is particularly suitable for the load securing function in the load holding brake valve. Since there is a downstream pressure relief valve in the directional control valve in the usual circuits, this does not lead to a summation of the set pressures.

FIGS. 5 to 10 show a possibility of hydraulic stroke limitation of a pilot operated valve. This hydraulic stroke limitation can be applied to all hydraulically pilot-controlled valves in which a pilot valve is provided for actuating a valve piston. The hydraulic stroke limitation is illustrated on a load holding brake valve, as described in FIGS. 1 to 4. 5 is similar to the switching piano according to FIG. 1. Reference is made in full to the description of FIGS. 1 to 4. 3, 4 is not shown here. The load holding brake valve is supplemented by a metering valve 84 in the control of the control piston 20 via the control connection X.

A metering valve 84 is used for this control. The metering valve 84 is shown in detail in FIG. 6 and is described with reference to FIG. 6. The metering valve 84 is located in the cover 22, which delimits the opening chamber 21. The cover 22 is flanged to the valve housing 1 of the load-holding brake valve in a pressure-tight manner by means of a seal 121. The metering chamber with valve seat 109 of the metering valve 84 is movably guided and positionable relative to the control chamber 21. For this purpose, e.g. B. the valve seat 109 of the metering valve 84 is formed on a locking piston 119, which closes the metering valve chamber 102 from the control chamber 21 in the rest and which is guided and positioned in the metering valve chamber 102 parallel to the control piston in a sealing manner. For this purpose, there is a longitudinal bore 104, 105 in the cover 22, which is coaxial to the valve axis of the load-holding brake valve. This longitudinal bore is provided with a thread 105 at its end which faces away from the load holding brake valve. On its remaining length (connection step 104) it has a larger diameter. An adjusting spindle 106 is screwed into the thread 105 and clamped pressure-tight with a counter sealing nut 113. The adjusting spindle 106 forms with the longitudinal bore 104,

105 an annular space in the area of the connection stage 104. The control line opens into this connection stage 104. A filter 116 and a nozzle 117 are switched into the control line X. The annular space is closed off on the side facing the load-holding brake valve by a guide collar 119 which is firmly connected to the end of the adjusting spindle 106 and which is in the guide step 103 of the longitudinal bore 102 by means of seals designed as an O-ring 120 is guided sealing. The adjusting spindle

106 is penetrated centrally by a central channel 108. At the end facing away from the load holding brake valve, the

Central duct 108 is closed in a pressure-tight manner by a stopper 12. At the end of the central channel 108, which faces the load-holding brake valve, the central channel 108 opens into the control chamber 21 with a valve opening channel 107. The metering valve with the closing element 110 and a shaft 118 is located in front of the valve oil channel 107 Closing element 110 is a ball here. The shaft 118 is supported on the one hand on the closing element 110 and is preferably firmly connected to the control piston 20. The shaft 118 penetrates the valve opening channel 107 with great play and projects into the Aufsieuerkammer 21, where it rests on the front side of the control piston 20, which delimits the control chamber 21 on the other side. The central channel 108, together with the smaller diameter opening channel 107, forms a conical or dome-shaped annular valve seat 109, on which the closing element 110 fits. The closing element 110 is guided with play in the central channel 108. It is pressed by a spring 111 in the direction of the control piston 20 in such a way that it is supported by the shaft 118 on the end face of the control piston 20. In the depressurized state of the control chamber 21, the control piston 20 bears against the cover 22 in which the metering valve is located under the force of the springs 46, 47. In this position, the shaft 118 supports the closing element 110 so far from the valve seat 109 that there is space for a radial channel 114 which connects the bore 102 and the control channel X opening into it via the connecting step 104 with the central channel 108.

About the function:

When the control port x is subjected to control pressure, the control pressure propagates in the connection stage 104 and the radial channel 114 into the central channel 108. Since the closing element 110 has a large play in relation to the walls of the central channel 108, the control pressure is on both sides of the closing element 110. The oil flow then passes through valve opening 107 into the opening space 21.

The closing element 110 and the shaft 118 of the closing element are pressed by the spring 111 in the direction of the control piston 20, so that the shaft 118 and the locking element 110 the opening movement of the opening participate in the control piston. In this case, the closing element 110, which is designed here as a ball, comes to the end of the central channel 108 and comes into contact with the valve seat 109 of the valve opening. As a result, the valve opening 107 is closed and the opening movement of the opening piston 20 is ended.

This state of the hydraulic stroke limitation of the control valve is shown in FIG. 7, to which the description of FIG. 6 also applies. The closing element 110 preferably lies on the seat 109 without leakage.

When the control connection is relieved of pressure, on the other hand, the control piston 20 moves under the load of the springs 46, 47 to its stop, ie. H. the lid 22 back.

It is in an emergency, i. h .: if the hydraulic control pressure fails, it is also possible to mechanically actuate the control piston 20. For this purpose, the adjusting spindle 106 is rotated in its thread 105 such that the front end of the adjusting spindle 106 - d. h .: the guide collar 119 strikes against the end face of the control piston 20 and moves the control piston in the direction of the main piston 3 in the sense of an opening of the pilot valve with pilot valve seat 6. This makes it possible to lower the load without control pressure. This operating state is shown in FIG. 8, to which the description of FIG. 6 also applies.

FIGS. 9 and 10 show a further embodiment of the metering valve for actuating the control piston 20. With regard to the description of the metering valve, reference is made to the description of FIGS. 5 to 8. In addition, the following three elements are shown here, which are only for can be used together or in combination with two or three with the metering valve:

a) Dosing bypass A dosing bypass duct 126 branches off from the annular duct 104, to which the control pressure can be applied, via a damping nozzle 125. This metering bypass channel 126 continues in a branch channel 127 which opens into the control chamber 21. If necessary, a further damping nozzle 128 can be arranged in the branch channel 127.

About the function:

Through the metering bypass channel 126 with one or more damping nozzles 127, the control chamber 21 is also acted upon by the control pressure when the closing element 110 closes the valve seat 109. However, there is only a damped to the desired extent

Actuation of the control piston 20. This also changes the function of the metering valve.

When the metering bypass is used, the metering valve effects an unhindered, rapid actuation of the control piston 20 in the first control area. The metering valve causes the main valve to respond quickly, ie the load holding brake valve in lowering mode. This quick control range is ended when the metering valve prevents the control oil from flowing in through valve opening 107 (hydraulic stroke limitation of the quick control range). Now the control chamber is only heavily throttled with control oil via the bypass. The acceleration reduction is slowed down accordingly. In this state, only the metering bypass 127 remains effective, so that the load holding brake valve can be operated sensitively. Without the use of the metering valve, the requirements would be: large control path with Faster control with good damping is only possible if the nozzles required for damping were inserted into control channel x and a very high control pressure was used to quickly control the control movement. By adjusting the adjustment spindle, the ratio of the quick control range to that

Total control range can be set.

On the other hand, the use of the less throttling metering valve also enables a rapid return movement of the control piston 20, since the two damping nozzles 125, 128 in the metering bypass 126 are bypassed through the valve opening 107.

b) tank bypass

A tank bypass channel 137 branches off from the metering bypass and connects the metering bypass to tank channel 138. A bypass nozzle 132 and a bypass check valve with ball 133 and spring 134 are arranged in the tank bypass channel 137. The check valve prevents the backflow in the metering bypass 126 from the leakage oil seal L via the nozzle 132 to the connecting bore.

About the function:

The pressure in the connecting bore 126 opens the ball 133 of the bypass check valve. As a result, part of the control oil flows through the bypass nozzle 132 and the bypass channel to the tank. This creates a flow and pressure division in the metering valve. This dampens pressure vibrations. The strength of the damping can be determined by the size of the bypass nozzle 132.

c) preloading the control pressure A pretensioning bypass 129, 131 branches off from the annular space 104, which is loaded with the control pressure. In this preload bypass is a preload valve (pressure relief valve 130) which can be adjusted by means of a screw. The preload valve has, as is known, a spring-loaded check valve which is opened by the pressure in the annular channel 104 and establishes a connection with the control chamber 21.

About the function:

If the control pressure in the annular space 104 and o in front of the damping nozzle 125 suddenly increases, the preload valve opens

130. The control oil thus flows quickly and directly into the control chamber 21. There is a rapid reaction of the control element in the sense of an opening of the load holding brake valve for lowering the load.

5 While rapid actuation is already made possible by the metering valve during normal operation of the load holding and braking valve, the preload valve used in combination with it enables further accelerated actuation bypassing the rapid control area and the fine control area. 0

It should be noted that the metering valve can be used on its own, but also in combination with one or more elements a, b and c, for other control tasks which involve a control piston through which a hydraulic current flows is controlled to be hydraulically controlled and adjusted by a control pressure, in particular to be adjusted against the force of a return spring. LIST OF REFERENCE NUMBERS

1 valve housing

1A load holding brake valve

2 control chamber

3 main pistons

4 sealing surface

5 valve seat

6 pilot valve seat

7 sealing surface

8 pilot spools

9 pilot tappets, piston rod

9A approach

10 throttle groove

11 spring plates

12 spring

12A spring

13 control chamber plug

14 compensating choke

15 spring chamber, pilot control chamber

16 Steering shaft, thin part

17 separator

18 seal

19 Taxation, big part

20 control pistons

21 control room

22 plugs

23 Mittelbund

24 control spring 25 lowering line

26 consumers

27 return line

28 inlet pipe

29 control line

30 pressure relief valve

31 way valve

32 pump

33 tank

34 pilot channel, seat bore

35 Undercut groove

36 throttling point

37 leadership

38 guide bore

39 end face

40 anteroom

41 Pre-throttle bore, pre-throttle

42 final waistband

43 control bore, spring chamber

44 end face

45 end face

46 spring

47 spring

48 stop surface

49 Overload hole

50 blind hole

51 stoppers

52 inner bore

53 radial bore

54 Check valve seat 55 pressure relief pistons

56 sealing surface

57 compression spring

58 piston attachment

59 bottom of hole

60 return hole

61 seal

62 final covenant

63 final covenant

64 end face

65 end face

66 compression spring

67 stoppers

68 hole

69 relief bore

70 annulus

71 stepped bore

72 level

73 return space

74 guide bore

75 end face

76 insertion

77 guide bore

78 insertion

79 seal

80 branch channel

81 relief channel

82 seal

83 lock nut

84 dosing valve 101 lid

102 longitudinal bore

103 management level

104 connection level

105 threads

106 adjusting spindle

107 valve opening channel

108 central channel

109 valve seat

110 closing element

111 spring

112 stoppers

113 counter seal nut

114 radial channel

115 control connection

116 filters

117 nozzle

118 shaft

119 locking piston

120 seal (O-ring)

121 seal (O-ring)

125 damping nozzle

126 connecting hole, hole

127 connecting bore, damping channel

128 damping nozzle

129 connecting bore, preload channel

130 preload valve

131 connecting channel, preload channel

132 bypass nozzle

133 Bypass check valve valve ball Spring of the bypass check valve Bypass channel, bore seal (O-ring) Bypass channel tank channel, connecting bore undercut groove Area with the largest cross section Area with decreasing throttling (throttling effect) Part length with decreasing cross section Part length with constant cross section Area with smallest cross section End chamber

Claims

PATENT CLAIMS

1. Hydraulically controllable load-holding brake valve, in particular for a double-acting consumer, which is on one side on its load side under an external load, with the following features: a control chamber (2) is arranged in a valve housing (1); the control chamber consists of, preferably aligned,

Chamber sections in this order:

Pilot room (15); Annulus (70) which (via connection B) with the lowering line (25) of

Consumer (26) is connected,

Return space (73) which is connected (via connection A) to the return line (27) to the tank;

Control chamber (21) which is connected to a control channel (X); A valve seat (5) with a central passage is arranged in the control chamber (2) between the annular space and the return space in the control chamber (2), via which the connection bores A and B can be connected; the valve seat is closed or opened by a main piston (3); the main piston (3) is designed as a stepped piston and has: a thin piston collar, which with the cylindrical wall of the

Control chamber (2) forms the annular space (70), a sealing surface (4) on the thin piston collar, which faces the valve seat and which cooperates with the valve seat (5), a thick piston collar, which seals against the wall of the

Control chamber is guided between the annular space and the pilot space and separates the two; the main piston is axially displaceable in the control chamber (2)

Pressurizing the return chamber (73) or the annular space (70) in the sense of lifting off the valve seat (4) and by means of pressurizing striking the pilot control chamber (15) in the sense of closing the valve seat; the pilot chamber (15) is connected to the annular chamber (70) and port B via a compensating throttle (14) and via a pilot channel (34) with pilot valve seat (6) in the main piston (3) to the return chamber

(73) and connection A can be connected; the pilot channel with pilot valve seat (6) can be closed by a concentric to the pilot channel (34) closing element (pilot piston 8) with its sealing surface (7) by applying pressure in the pilot chamber (15) and, preferably, the force of a closing spring (12) and in the The opposite direction can be unlocked by a pilot tappet (piston shaft (9)), which pilot tappet (9) is guided with play in the pilot channel (34) and projects into the return chamber (73); A pilot piston (20) is guided axially in the pilot chamber (21) and can be displaced in the opposite direction by pressurizing the pilot chamber (21) in the direction of the return chamber (73) and by a pilot spring (24); the control piston (20) has a control shaft (19) directed against the pilot tappet (9) and coaxially aligned with it, which projects with its one end (control end (16)) into the control chamber (2) and upon axial displacement of the control piston (20 ) acts against the force of the control spring (24) on the pilot tappet (9) and the pilot piston (8) in the sense of opening, characterized in that the pilot tappet (9) over its length and starting from the seat surface (7) of the Pilot piston (9) has at least the following length ranges: first, an area (142) with a maximum cross-section, which is guided with minimal play (throttle gap) relative to the pilot channel (34), then an adjoining throttle area (143), which forms a throttle gap over its length with its cross section relative to the pilot channel (34), which starts from the throttle gap of the maximum cross section and then, at least over a partial length (144) of the throttle area (143 ), increasing continuously, preferably progressively, then an area (146) with a minimum cross section; and that preferably the pilot tappet (9) is firmly connected to the pilot piston (8).

O

2. Valve according to claim 1, characterized in that the pilot tappet (9) in the throttle area which is closest to the pilot piston (area 142 of the maximum cross-section), circular-cylindrical and essentially with the maximum cross-section and with little play relative to the pilot channel 5 (34) is designed so that the pilot tappet (9) in the then following "throttle area (143) with decreasing throttle effect" has at least one axially directed throttle groove (10) on its surface, the depth and / or width of which corresponds to the area of Maximum cross-section connects with essentially 0 zero and increases continuously over a partial length (144) of the throttle area (143), and - preferably - then continues over a further partial length (145), and - preferably - that the groove base of the throttle grooves on the other End of the throttle area (143) essentially runs out on the minimum cross-section of the pilot tappet (9).

3. Valve according to claim 1 or 2, characterized in that the effective surface with the control pressure (45) of the control piston 0 (20) to the active surface of the pilot valve seat (6) in relation of greater than 50: 1, preferably greater than 100: 1, and that, preferably, the ratio of the end face (45) of the control piston (20) to the end face (44) or effective area (74) at the control end (16) is greater than 30: 1 , especially larger than 60: 1.

4. Valve according to one of claims 1 to 3, characterized in that the throttle cross-section (throttle point 36) which the pilot tappet (9) forms with the pilot channel (34) is smaller in all opening positions of the pilot piston (8) than that between the Pilot valve seat

(6) and the sealing surface (7) of the pilot piston (8) formed opening cross section.

5. Valve according to claim 4, characterized in that the maximum throttle cross section which the pilot piston (8) forms with the pilot channel (34) is greater than the flow cross section of the compensating throttle (14).

6. Valve according to one of the preceding claims, characterized in that the main piston (3) at its end facing the pilot chamber (15) has a central guide bore (38), from the bottom of which the pilot channel (34) extends ; the pilot piston (8) on the pilot chamber facing

(spring-loaded) end has a guide shaft (37) which is sealingly guided in a guide bore (38) in the main piston (3) and has a larger end face (39) compared to the active surface of the pilot valve seat (6); the part of the guide bore (38) which lies between the pilot valve seat (6) and the guide shaft (37) is connected to the pilot chamber (15) via a pilot throttle bore (41).

7. Valve according to one of claims 1 to 6, characterized in that the annular space (70) (including connection B and sink line 25) and the return chamber (60) with the return space (73) (including connection A, return line (27) and Tank) via a chamber (49) and a return chamber (60) and a spring-loaded pressure-limiting piston (55) of a pressure-limiting valve (30) arranged between them.

8. Valve according to claim 7, characterized in that the chamber (49) and the return chamber (60) between two Endkam¬ mers of the pressure relief valve (30) that the spring-loaded pressure relief piston (55) of the pressure limiting valve (30) has a sealing surface (56) and at both ends each have a piston shaft (guide end 62, guide end 63), the sealing surface (56) abutting the valve seat (54) under the biasing force of a compression spring (57) and each guide end (62, 63) in one of the end chambers of the valve housing (1) is sealingly guided so that the end chamber which is adjacent to the overload bore (49) and its leading end (62) with the pressure of the return chamber (73) via the longitudinal bore (81) and transverse bore (80 ) loaded and in cross section (end face 64) is slightly smaller than the seat surface of the valve seat and that the end chamber with the guide end (63) is relieved of pressure and with regard to its hydraulically effective cross section (end face 65) is the same size as the difference between the valve seat surface (54) and the cross section (end face 64) of the end chamber (147) with the guide end (62 ).

9. Valve according to claim 7 or 8, characterized in that the spring-loaded pressure-limiting piston (55) of the pressure-limiting valve (30) is loaded with two pressure springs connected in parallel, one of which is a compression spring (57) in the return chamber (60) against the Pressure limiting piston (58) and the other pressure spring (66) is clamped in the pressure-relieved end chamber against the piston shaft with the guide end (63).

10. Valve with hydraulic / mechanical control piston which can be actuated by an adjustable control oil flow in the sense of opening or closing the valve, in particular valve according to the preamble of claim 1, preferably according to one of the preceding claims, characterized in that the Control chamber (21) is connected to the control channel (X) via a metering valve (84) through which the control piston (20) is acted upon by a control oil quantity limited to a predetermined stroke of the control piston (20).

11. Valve according to claim 10, characterized in that the metering chamber (102) of the metering valve (84) is connected to the control connection (115) and has: a valve opening (107) with a valve seat (109) through which the control oil reaches the pilot chamber (21), a closing element (110) which is supported on the pilot piston (20) by means of a shaft (118) in such a way that the closing element ( 110) in the metering chamber (102) between the valve seat (109) and one

Opening position can be moved synchronously with the control piston (20) and that it closes the valve seat (109) at a predetermined stroke of the control piston (20).

12. Valve according to claim 11, characterized in that the valve opening (107) opens into the control chamber (21) on the side facing away from the control spring (147) and is surrounded by an annular closing surface (valve seat 109) which is parallel to pressurized end face of the control piston, the shaft (118) penetrates the valve opening (107) with great play, the closing element (110) by means of a spring (111) for bearing the shaft against the pressurized end face of the control piston (20) and after passing through it of the specified stroke is pressed against the valve seat (109).

13. Valve according to claim 12, characterized in that the valve seat (109) with valve opening (107) of the metering valve (84) relative to the control chamber (21) is movably guided and positionable, in particular the valve opening (107) of the metering valve ( 84) is formed on a closure piston (119) which closes the metering valve chamber (102) with respect to the control chamber (21) and which is sealingly guided and positionable in the metering valve chamber (102) parallel to the control piston (20).

14. Valve according to claim 13, characterized in that the closure piston (119) can be positioned such that it abuts the control piston (20) and the control piston (20) in the sense of unlocking the pilot closing element (pilot piston 8) moves and positions .

15. Valve according to claim 13 or 14, characterized in that the locking piston (119) is attached to the free end of an adjusting spindle (106), the adjusting spindle (106) has a central channel (108) aligned with the valve opening (107), which at the free end of the adjusting spindle (106) is closed by the plug (112), in the central channel (108) the closing element (ball 110) is guided, the central channel (108) is acted upon on both sides of the closing element with the control pressure and the adjusting spindle (106) can be screwed into or out of a threaded bore (105) parallel to the movement of the control piston (20).

16. Valve according to claim 15, characterized in that in the one end position of the adjusting spindle (106) the locking piston (119) projects into the control chamber (21), strikes the control piston and the control piston (20) in the sense of unlocking of the pilot control closing element (pilot piston 8) moves (Fig. 8) and in the other end position the distance of the valve seat (109) from the end face of the control piston (20) which is in the rest position is shorter than the shaft (118).

17. Valve according to claim 15 or 16, characterized in that the central channel (108) on both sides of the closing element (110) is acted upon by the control pressure by the control pressure channel (114) immediately in front of the closing surface of the sealing piston in the central channel

(108) opens and the closing element is guided with play in the central channel.

18. Valve according to claim 10 to 17, characterized in that the shaft (118) is fixedly connected to the closing element (110), or is separate from the closing element (110); the shaft (118) is fixedly connected to the control piston (20), or is separate from the control piston (20).

19. Valve according to one of claims 10 to 18, characterized in that the metering valve (84) is bypassed by a throttle channel (127) which, after closing the seat (109) by the closing element (HO), a stronger throttling (throttling or Apertures 125 and 128) of the control oil flow.

20. Valve according to one of claims 10 to 19, characterized in that the metering valve is bypassed by a biasing channel (129, 131) with a biasing valve (130) therein, with the maximum pressure difference between the biasing channel (129) and residual control channel (21) is specified.

21. Valve according to one of claims 10 to 20, characterized in that the metering valve is bypassed by a relief channel (bypass channels 135, 137) which connects the control channel to the tank via a bypass nozzle and a check valve 133).