EP0883753B1 - Load-holding brake valve - Google Patents

Description

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 as a pilot valve a ball seat valve, which is very sensitive control by the pilot piston required to one of Starting to allow steady flow increases and a leaping Avoid opening behavior.

It is therefore the object of the invention to provide a load holding brake valve of the o. G. 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 A progressive opening behavior of the valve is achieved, the dismantling of the load volume flow take place very evenly and steadily.

A progressive opening behavior is understood here to mean that the Opening is essentially proportional to the control pressure, anyway but a defined, jerk-free and shock-free dependency between the opening cross-section and control pressure (i.e. the 1st and 2nd derivative of the function: "Opening the valve above the control pressure" are for any change in the Control pressure finite and steady).

This task is accomplished by a load holding brake valve with the features solved according to claim 1.

Here, the pilot piston has one following its sealing surface Piston shaft (pilot tappet) with little play in the seat bore to be led. The piston skirt has a series of several over its length Cross-sectional areas. The maximum cross section follows and corresponds to that Seat section. In essence, it has very little play 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. It forms the valve lifter a throttle point in the seat bore (pilot channel), its throttling effect when moving the plunger and / or the resulting appearance the pilot channel steadily - and preferably progressively - becomes smaller. At one point in the throttle area, preferably shortly after If the pilot spool is lifted from its seat, the throttling effect is so large that it is much larger than the throttling effect of the compensating throttle. From there the throttling effect takes place in relation to the pilot channel with increasing control path of the control piston and displacement of the pilot tappet in such a way that the function of the throttle effect is constant over the displacement length of the pilot ram. Preferably takes the throttling effect initially slight and then with increasing Displacement path increasingly; The length and cross section of the ram are So matched so that the plunger with the pilot channel of the main piston a very small throttle gap at the start of opening of the pilot piston forms, the throttling effect much greater than the throttling effect is the compensating throttle and then until the minimum cross-section is reached forms a steadily increasing throttle gap, the Throttling effect with the movement length of the pilot tappet and Control piston steadily and preferably decreases more and more, so that the Remove the closing forces acting on the main piston. This is through the special design of the throttle area with regard to its length and reached its cross section. In particular, there are two cross-sectional designs possible:

In the first embodiment, the cross section starts from the maximum cross section and then takes over its length to the minimum cross section steadily. Here, the decreasing throttling effect achieves that the throttle gap of the pilot tappet relative to the pilot channel wall, starting from the throttle gap of the maximum cross section, steadily increasing.

In the second embodiment, the cross section also starts from the maximum cross section out; he then takes over a first part length except for one Cross-section that is larger than the minimum cross-section; about another This cross section remains the same for part of the length. Here the decreasing Throttling effect achieved in that the immersed in the pilot channel Length of the throttle gap of the pilot tappet with displacement of the pilot tappet is steadily decreasing. Combinations of both versions are also conceivable.

The constant reduction in the cross section of the tappet in the throttle area d. h .: the decrease in the throttling effect of the tappet compared to the pilot channel can e.g. can be achieved in that the plunger in the area with a reduced cross-section is a rotary body whose diameter changes Slightly conical over its length or - preferably - progressive, i.e. parabolically or hyperbolically reduced. It would also be possible to Area with a decreasing cross section cylindrical with the Form the diameter of the area with the maximum cross section in which Area of the decreasing cross-section, however, a chamfer or Flattening or axial throttling grooves of variable depth and width to be provided starting from the area of maximum cross section and to the circumference of the area with the smallest cross-section. to Achieving a progressive reduction in cross-section can mean depth additionally or alternatively, however, the width of these chamfers or Increase throttling grooves parabolically or hyperbolically.

The length of the throttle region is preferably based on the change in the the closing piston acting on the main piston. These closing forces are composed of the hydraulic forces and those on the Main piston acting spring forces.

This adjustment is done in such a way that the tappet with the pilot channel of the main spool at the start of opening of the pilot spool small throttle gap forms, the throttling effect greater than the throttling effect is the compensating throttle and then until the minimum cross-section is reached forms a steadily increasing throttle gap, the Throttling effect with the length of movement of the pilot tappet and pilot piston decreases more than those acting on the control piston Closing forces increase. In the design of the valve with a direct the spring acting on the main piston is the length of the throttle area inversely proportional to the spring stiffness on the main piston Closing springs. Through this training it is achieved that after opening the pilot valve the pressure reduction in the pilot room starts slowly and depending on the length of the plunger and the Movement length of the control piston or pilot ram 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 it is the plunger and its throttle area. In other words: the stronger the Spring force on the main piston increases when opening, the stronger must when the pilot tappet moves, its throttling effect in the Remove the pilot channel.

This results in a stable equilibrium position with each opening pressure of the main piston. This also makes a sudden, jerky and erratic Prevents opening or moving of the load holding brake valve piston. The Opening characteristics and adjustment of the ram length on the one hand and the On the other hand, springs also prevent the excitation of vibrations the moving loads.

The pressure reduction in the pilot control room takes place continuously and in a defined manner Dependence on the control pressure and the resulting movement of the The opening piston. It is avoided that the main piston is the pilot piston hurries ahead or uncontrolled and uncontrollable movements performs.

A precisely working hydraulic follow-up system is created. Once the Load pressure is released behind the main piston, the main piston follows automatically the pilot piston, since the load pressure on the resulting Annular surface of the main piston acts, the main piston from its valve seat pushes. Because the main piston is the pilot piston displaced by the pilot piston follows, the opening cross-section becomes in the pilot valve seat again narrow so that there is again a back pressure in the pilot room of the main piston. This creates an equilibrium position between the control piston and the pilot piston on the one hand and the main piston on the other hand. The advantage of this principle is that the closing direction acting flow force through the hydraulic power amplification achieved is less than the hydraulic opening force under all conditions. This prevents pressure fluctuations in the consumer connection B trigger unwanted movements of the main piston. This can do that A boom on the hydraulic excavator or crane is prevented from rocking become.

The maximum cross section of the pilot tappet is relative to that Cross-section of the pilot channel designed so that the throttle cross-section after opening the pilot valve seat is initially only insignificantly smaller than the throttle cross section of the compensating throttle, so that when lifting the Main piston from the seat a slow pressure reduction in the pilot room can take place. By designing the area with maximum cross-section and the adjoining area with a reduced cross section Different 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 is.

Overall, the load holding brake valve according to the invention integrates the functions for holding, lowering, lifting and securing loads in one valve housing with a very compact design. It is achieved that at accordingly designed course of the throttle cross-section constantly a hydraulic Force acts on the pilot piston so that this or the tappet itself cannot lift off the control piston, despite the pressure drop on the Spring side of the pilot piston.

In the load lifting state, the spring-loaded main piston also takes over Function of a check valve. Here, the small opening pressure of the Check valve made possible by a large seat. By the big one Area ratio between the effective diameter of the valve seat and the Effective diameter of the pilot valve seat is reached that the pilot piston does not open.

The invention allows a fine gradation of the various throttling points, which are formed or are present in the valve. In the design of the valve according to claim 1, the control pressure is largely independent of the load pressure to be controlled, so that even with low Control pressure allows a wide control range and sensitive control becomes.

By training according to claim 1 it is achieved that the opening of the Valve seat, which due to the size of the valve seat diameter only one inaccurately definable throttling does not have a negative impact on the Has opening behavior of the valve.

The training according to claim 1 ensures that the control with a low control pressure, the valve reacts and a Movement of the load leads.

According to the invention, it is possible to open and close the Guide the main piston hydraulically alone, as on both sides of the main piston always definable by fine-tuning the throttling points hydraulic conditions prevail.

The embodiment according to claim 1 serves to increase the closing effect. for safety two springs are preferably provided in parallel can be. One of these springs can only be used on the main spool act, but the other on the pilot spool and thus indirectly also on the main piston. It is advantageous because of the sensitive pilot control of the main piston also possible, only the pilot piston to load with a spring in the closing direction, which at the same time also acts on the main piston in the closing direction.

Regarding the formation of the pre-throttle bore in the guide shaft there is extensive freedom of design, 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 room and annulus.

The embodiment according to claim 3 is advantageous if the control pressure is specified regardless of the inlet pressure. By grading the Pilot piston and the pilot throttle is brought about that a higher closing pressure acts on the pilot piston. In particular In open systems with external control, the flow can be reduced be reduced from B to A with increasing load pressure. The reaction of the The valve is damped so that there are discontinuities or vibrations Stress or discontinuities in the control in particular not as Vibrations of the valve can affect.

Because the pilot piston and the pilot piston are independent misaligned, misalignments remain between the control piston and the pilot spool with no effect.

In the load holding brake valve according to claim 4 is in the valve housing Pressure relief valve integrated to protect the load pressure. Moreover the highest load pressure can be set easily.

It is advisable to take measures to reduce the spring load to reduce the pressure relief piston. This will make this goal ensures that the pressure limiting piston has only a small effective area, which is loaded by the load pressure. This will provide the necessary spring force greatly reduced and the installation space reduced.

Such an embodiment results from claim 5.

For safety reasons, two grip the pressure limiting piston Compression springs, which are connected in parallel, but to save space are arranged according to claim 6.

When executing the load holding brake valve according to one of claims 4 up to 6 the opening pressure of the pressure relief valve is also independent from the return pressure. This principle of the pressure relief valve is achieved that with a downstream pressure relief valve in the directional control valve there is no summation of the set pressures.

The load holding and braking valve according to this invention has - as already executed - a control piston with which the actuation of the pilot piston hydraulically and mechanically. Such hydraulic-mechanical actuation of valves comes in hydraulics many times before, e.g. B. in the hydraulic actuation of control valves. This hydraulic control has the disadvantage that the opening of the Control valve for a - more or less fast - growth of the Control oil quantity leads. It therefore depends on the attention and Dexterity of the operator from that the control valve on reaching a desired drive oil quantity, d. H. when a certain one is reached Position of the hydraulic-mechanically controlled valve, the To keep the control valve in this position.

Another object of this invention is this, like any other hydraulic-mechanical controlled valve, in a predetermined end position bring. Such an embodiment results from claim 7.

Such metering valves can be based on a hydraulic principle of action, z. B. by metering predetermined amounts of oil, which then as the control oil Control piston are supplied.

However, this would require the amount of oil to be metered in each case adjust the desired path of the control piston. This adjustment happens automatically when executing the valve according to claim 8.

Here, the closing element can be mechanically connected to the Control piston be contacted so that when reaching a predetermined Position of the control piston the supply of another control oil flow is interrupted.

A hydraulically and mechanically advantageous integration of the metering valve in the control device results from claim 9. In this embodiment there is an easily accessible and easy-to-use option of Setting the dosing amount from claim 10. Furthermore, here is one Emergency actuation of the control piston desired and achieved through the design according to claim 11, 12 and 13. It is on the one hand mechanical actuation of the control piston in the event of control pressure failure and on the other hand a complete deactivation of the controllability of the control piston causes. Both functions can be used for security reasons be required.

Claim 14 gives a simple way of relieving the pressure Closing element.

Claim 15 describes further execution options.

These designs thus on the one hand limit the stroke Control and on the other hand the possibility of mechanical unlocking, especially 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 represent only one Position from which the end position is to be approached. This is especially when lowering loads. There is said to be the essential The route is traversed at high speed, while the end position is slow, d. H. is started in slow speed.

Claim 16 forms the valve according to claims 7 to 15 accordingly out. With this training, the fast mode can be switched on very suddenly the valve in the open state of the metering valve is operated. After the end stroke of the metering valve has been reached on the other hand, an operation in creep speed via damping nozzles, which is an adjustment allow the speed of the creep speed. The now damped operation of the control piston allows the desired end position to start up by fine control. The ratio of the overdrive range to slow speed (fine control range) can be set from the outside by adjusting the Adjustment spindle for the metering valve can be set. The quick response of the control piston remains possible despite the strong hydraulic damping. In order to have a outside of the functional range of the metering valve To enable quick reaction, a preload valve is provided, which when a certain control pressure is exceeded, the connection of the control pressure channel to the control chamber and control piston opens (claim 17).

The training according to claim 18 serves the purpose of pressure vibrations the control pressure both in fast mode and in fine control mode dampen.

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.

Fig. 1

Hydraulischer Schaltplan zur Steuerung eines Verbrauchers im Sinne einer Anpassung eines Ablaufstroms an einen Zulaufstrom,

Fig. 2

einen Längsschnitt einer Variante eines Lasthalte-Bremsventils,

Fig. 3a/b

einen Längsschnitt durch ein Vorsteuerventil,

Fig. 4

einen Längsschnitt einer Variante eines Lasthalte-Bremsventils,

Fig. 5

hydraulischer Schaltplan entsprechend Fig. 1 mit hydromechanischer Begrenzung des Aufsteuerstroms,

Fig. 6

Detail des Dosierventils nach Fig. 5,

Fig. 7

das Detail nach Fig. 6, jedoch im Zustand des ausgefahrenen Hub des Ansteuerkolbens, d. h.: bei hydraulischer Hubbegrenzung

Fig. 8

Detail nach Fig. 5, jedoch mit mechanischer Aufsteuerung des Aufsteuerkolbens,

Fig. 9

hydraulischer Schaltplan entsprechend Fig. 1 mit einer hydraulischen Dämpfung des Ablaufstroms, mit Überdruckventil, Dämpfungsbypass sowie Entlastungs-Bypass.

Fig. 10

Ausführungsbeispiel nach Fig. 9.

Show it:

Fig. 1

Hydraulic circuit diagram for controlling a consumer in the sense of adapting an outflow to an inflow,

Fig. 2

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

3a / b

a longitudinal section through a pilot valve,

Fig. 4

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

Fig. 5

1 with hydraulic limitation of the pilot current,

Fig. 6

Detail of the metering valve according to Fig. 5,

Fig. 7

the detail of FIG. 6, but in the state of the extended stroke of the control piston, ie: with hydraulic stroke limitation

Fig. 8

5, but with mechanical actuation of the actuating piston,

Fig. 9

1 with hydraulic damping of the discharge flow, with pressure relief valve, damping bypass and relief bypass.

Fig. 10

Embodiment according to FIG. 9.

In Fig. 1 is the hydraulic circuit diagram of a control of a consumer in the sense of adapting an outflow to an inflow by means of a load holding brake valve shown. The consumer 26 is on the feed line 28 and the lower line 25 connected. The lower line 25 is connected to port B of the load holding brake valve 1A. From the load holding brake valve 1A leads a return line 27 from port A to the Directional control valve 31. The inlet line 28 also connects to the directional control valve 31. The Directional valve 31 is designed as a 4-3-way valve. In addition to the inlet pipe 28 and the return line 27 is the connection of a pump 32 and the connection of a line to the tank 33 is provided. The load holding brake valve 1A is connected to the feed line 28 via a control line 29. Furthermore, there is between the lower line 25 and the return line 27 the pressure relief valve 30 switched. In the switch position shown is the inlet line 28 and the return line 27 with the tank 33 connected. Thus, the consumer 26 remains in the current one Position. The connection between port B and port A of the load holding brake valve 1A is locked.

If the slide directional valve 31 is shifted to the right, the supply line 28 connected to the pump 32. The consumer 26 is located now in lowering mode. Therefore, the return line 27 with the tank 33 connected. The connection between port B and the port A in the load holding brake valve 1A remains closed until the Pressure has built up in the inlet and is sufficient via the control line 29 Control pressure is present at the load holding brake valve 1A. Then it will Load-holding brake valve 1A moved to the right against the spring. Connection B and port A in the load holding brake valve 1A are now variable Throttle connected together. 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 for 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 moved to the left, the pump 32 connected to the return line 27. The inlet line 28 is also available the tank 33 in communication so that on the control side of the load holding brake valve 1A no pressure is applied and the load holding brake valve 1A in the position shown persists. The consumer is in this position 26 in lifting operation. The pump volume flow passes through the Return line 27 and the check valve in the load holding brake valve 1A to connection B. From there, the oil flows through the lowering line 25 to Consumer 26.

The pressure relief valve 30 serves to secure the load pressure in the Lowering operation or standstill of the consumer and is between the lowering line 25 and the return line 27 are arranged. An additional pressure protection is usually arranged on the directional control valve (not shown here).

Vorsteuerraum 15;

Ringraum 70, der (über Anschluß B) mit der Senkleitung 25 des Verbrauchers 26 verbunden ist,

Rücklaufraum 73, der (über Anschluß A) mit der Rücklaufleitung 27 zum Tank verbunden ist;

Aufsteuerraum 21, die mit einem Steuerkanal X verbunden ist;

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 control room 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 at the end by means of a control chamber plug 13 closed. In the control chamber 2 open perpendicular to the longitudinal axis the control chamber 2, the connection holes A and B. Between the Connection bores A and B have the control chamber 2 a valve seat 5 on. The valve seat 5 is fixedly attached to the valve housing 1 and separates the annular space 70 from the return space 73. Between one end of the Control chamber 2 with control chamber plug 13 and the valve seat 5 is a Main piston 3 guided. The main piston 3 has a thinner one Collar with a conical sealing surface 4, which cooperates with the valve seat 5. On the side facing away from the valve seat 5 and the connection bore B facing side, the main piston 3 has an end collar 42 on. The end collar 42 has a larger diameter than the above Bund and is sealingly guided in the control chamber 2, so that the main piston 3 is axially movable. By training as a stepped piston the main piston 3, the annular space 70, with the lowering line 25 via connection B is connected. The annular space 70 is by lifting the main piston 3 of the valve seat 5 with the return chamber, the connection B and the tank 33 connected.

The area of the control chamber 2 between the thick end collar 42 of the The main piston 3 and the control chamber plug 13 is the pilot control chamber 15 designated. This pilot control chamber 15 serves to receive a spring 12A (not shown) between the control chamber plug 13 and the main piston 3 is clamped. One is shown - to be described later Spring 12 - which has the same function so far, so that the main piston 3rd by spring force, but also by the hydraulic acting on it Forces on valve seat 5 is pressed.

The annular space 70 is connected to the pilot space 15 via the throttle 14. The throttle 14 can - as shown - axially parallel in the thicker Piston collar can be arranged, but also in the valve housing. The The main piston 3 is penetrated concentrically by a pilot channel 34, that connects the pilot room 15 to the return chamber 73. For this the main piston 3 has a concentric with respect to the pilot chamber 15 arranged step bore 71. From the bottom 72 of the first step with Larger diameter goes with the stage designated as pilot channel 34 smaller diameter. On the bottom 72 between level 71 and the pilot valve seat 6, the pilot valve seat 6 is formed.

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

The pilot tappet 9 has several diameter ranges over its length or cross-sectional areas.

A small groove adjoins the conical seat as an undercut. The Groove runs in the circumferential direction and has essentially manufacturing technology Reasons.

A very short area of large cross section (= Cross-sectional area) of the pilot piston 8. This area is cylindrical trained and with little play has a diameter that the diameter of the pilot channel 34 and the smaller sealing surface 7 corresponds. Its length can go to zero, so that it is only the beginning of the represents the following range.

An area is attached to the very short area of large cross-section with decreasing throttling effect. The one that decreases with the tappet movement Throttling effect is achieved in that the cross section of this Area - based on the maximum cross-section - at least over one Partial length steadily decreased and / or that the piece of this partial length that in the pilot channel is immersed, when the pilot tappet is moved shortened. Another part of this range can be constant Have a cross section, which is however larger than the cross section of the then following area with the smallest cross-section. The decreasing throttling effect results from the fact that with the tappet movement the area first with a decreasing cross-section from the pilot channel to the pilot room shows up. With further tappet movement, this changes into the pilot channel 34 immersed partial length of the pilot tappet 9, which is constant Cross section. The change in throttle effect in this area of the pilot plunger 9 is thus done by changing the throttle cross section and / or the throttle length immersed in and in the pilot channel 34 is led. That means the area with decreasing cross section or decreasing throttle effect no longer than the pilot channel 34 may be. The length depends in particular on the desired one Opening behavior in relation to the control pressure.

However, the area can be cylindrical with the diameter of the previous one Form the area and chamfer or on the cylinder jacket Make grooves that start from the largest cross section and on the smaller, constant cross-section. Fluidic and Production-technically favorable versions of the area with decreasing Cross sections are described with reference to Figures 3a and 3b.

The end of the pilot tappet 9 (piston skirt) has a smallest Cross section, which is essentially the smallest cross section of the area corresponds to decreasing cross-section. This end range is only too a part within the pilot channel 34. It extends over the length of the Pilot channel 34 and protrudes with its end in the return chamber 73 of the control chamber 2.

In particular, it is shown: The pilot tappet 9 points after the Sealing surface 7 on a circumferential undercut groove 35. That includes itself cylindrical area, whose diameter with play the diameter of the pilot channel corresponds to (area with the largest cross section, area with Maximum cross section).

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

The transition between the area requires special attention Maximum cross section and the area with decreasing cross section. This Transition must be continuous so that when the pilot tappet moves 9 in there is no impact or jerk in the load movement in this area.

In Fig. 3a, the decreasing throttling effect of area 143 on the first part length 144 through a decreasing cross section of the ram reached. The ram is weakly tapered on this part length, i. neck Truncated cone. The large cone area corresponds to the cross section of the previous area with the largest cross section. The small cone area corresponds to the cross section of the subsequent partial length 145, the one has a constant cross-section. This partial length 145 calls in the pilot channel only a small throttling effect that emerges with the appearance this part length decreases steadily from the pilot channel. This part length is therefore only of minor importance for the function of the valve. Their length can therefore go to zero. The design and is important Length of the previous area with decreasing cross section. It must to the drawing that the transition between the Area with maximum cross section and the area with decreasing cross section cannot be adequately represented there. In reality, may no circumferential edge is formed there, since a continuous, i.e. H. repair or hyperbolic transition is desired. Likewise, the parabolic or Hyperbolic design of the surface line cannot be represented. Is shown a linear surface line, which, however, is not considered preferred in the sense of this Invention can apply.

Partial length 145 with a constant cross section is included Area 146 with the smallest cross section. It should be emphasized that this is the smallest Cross section is in any case smaller than the cross section of the previous one Part length 145 with constant cross section. The boundary between the two Cross-sectional areas is in any case with the pilot valve closed the pilot channel. The area with the smallest cross section protrudes from the Pilot channel out into the return space.

The embodiment of the pilot tappet 9 shown in FIG. 3b has in the Area with decreasing throttle effect in the axial direction several throttle grooves 10 on that with the wall of the pilot channel 34, the throttle point 36 form. The throttle grooves 10 have a decreasing range Cross-section steadily towards the free end of the pilot tappet 9 and - preferably progressive - increasing depth (partial length with 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) follows here also the area with the smallest cross-section. This area is again cylindrical. The diameter can essentially that Correspond to the diameter of the deepest groove base of the throttle grooves 10.

The throttle grooves 10 can by flats or notches that are introduced axially or helically on the pilot tappet 9, replaced become. Instead of or next to the depth, the width of the throttle grooves 10 to be changed. This is especially true in the initial area of the grooves, i. h .: in the area with decreasing throttling effect. The grooves go from that Area with maximum cross-section with depth = 0 and width = 0. By increasing the width and depth of the grooves, a steady parabolic or hyperbolic or other courses of the cross section of the Reach the tappet.

About the function:

When the pilot piston 8 is axially displaced to the right, the pilot valve tappet opens 9 by the sealing surface 7 from the pilot valve seat 6 takes off. As long as the area with the largest cross section in the pilot channel 34 immersed (throttle point 36), the volume flow remains strongly throttled, whereby this throttling effect compared to the pilot throttle 14 the pressure reduction in the pilot control room and thus the opening behavior of the main piston certainly.

With increasing axial displacement of the pilot tappet 9 the Area of the largest tappet cross-section from the pilot channel 34 on and therefore steadily reduces its throttling effect. The throttling effect is now from the area with decreasing throttling, i.e. h .: first by the decreasing cross section of the pilot tappet emerging from the pilot channel 34 9 determined. Here the depth of the throttle grooves increases (Fig. 3b) or the diameter of the ram (Fig. 3a). The throttling effect When this part length (truncated cone or grooves) appears, the Pilot channel 34 in the pilot room steadily lower. If 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 continues However, the throttling effect continues to be significantly less, since the in length of the section immersed in the pilot channel with the same length Cross section decreases. That the area with the smallest cross-section remains submerged in the pilot channel 34 has no effect, because the throttling effect of this area is very low.

There is therefore a continuously slow pressure reduction. Here is the Opening cross section of the pilot valve seat 6 also immediately after Opening larger than the throttle cross-section in the pilot channel (throttle point 36).

A separating web 17 (FIG. 4) separates the return space 73 from that axially aligned control bore 43. The control bore 43 is on the other End face closed by the plug 22. In the control bore 43 is a control piston 20 (guide collar) sealingly guided. The control piston 20 divides the control bore 43 into the control chamber 21 and the the separator 17 adjacent spring chamber. The plug 22 has a connection bore X on, through which the control chamber 21 with the control line 29 (Fig. 1) is connected.

The control piston 20 has a control shaft 16, 19 which consists of a thicker part 19 and consists of a thinner part 16. The thinner part 16 of the control shaft penetrates the separating web 17 and is in the Separator in the guide bore 74 sealing (seal 18 or a sealing gap) guided. 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 tappet 9 of the pilot piston 8 lie on an axis line. The control piston 20 is formed by a control spring designed as a compression spring 24, which is arranged in the spring chamber 43 and on the Supported divider, pressed into its starting position when no control pressure rests in the control chamber 21. The spring chamber 43 is by means of Leakage oil hole L relieved of pressure. For safety reasons, the control spring 24 by one or more springs 46, 47 connected in parallel (see Fig. 4) formed.

The thicker area 19 of the control shaft 19 forms compared to thinner area 16 an end face 48. This end face serves as a stop surface 48 for mechanical stroke limitation of the control piston 20, by reaching the system with the separating web 17. For dimensioning be said that the guide collar of the control piston 20 one with one Control pressure acted on end face 45, the effective surface for Effective area of the pilot valve seat 6 in a ratio of greater than 50: 1, preferably greater than 100: 1.

Furthermore, that the ratio of the end face 45 on the guide collar to the end face 44 at the control 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 area of the control spring 24 thus enables a precisely predetermined and from Structure of the control pressure-dependent force curve on the control piston 20 acts. For safety reasons, it is advantageous if several springs act on the control piston 20. This ensures that even in the event of breakage a spring of the control piston 20 z. B. still controllable in the event of a line break is moved into its closed position.

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

The ratio of the effective areas of the main piston 3 and pilot piston 8 is designed in such a way 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 is executable.

About the function of the load holding brake valve:

At standstill:

In the connection bore B and the annular space 70 is the load pressure of the Consumer. The pilot control chamber 15 is connected to the throttle 14 Annulus 70 connected. The load pressure acts on the effective area of the thicker one End collar 42 of the main piston 3. The main piston 3 with its Sealing surface 4 is by the spring 12 and hydraulically against the valve seat 5 pressed.

The pilot piston 8 is with the load pressure and the spring force of the spring 12 acted upon; it is with its sealing surface 7 on the pilot valve seat 6 held. 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 with the pump and via return line 27 with the tank. The load holding brake valve is via the control line 29 and connection hole X via Inlet 28 connected to the pump. The one that can be changed by the directional valve Pressure acts as a control pressure on the control piston 20. The control pressure accordingly, the control piston 20 against the control spring 24th moved to the separating web 17 until the spring force and opening force in Are balance. The control shaft 16 abuts with its end face 44 to the free end of the pilot plunger 9 of the pilot piston 8 and shifts the pilot ram 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 will make the Connection between the return chamber 73 and the pilot chamber 15 created, the throttling effect of the design of the pilot valve 9th and the length of the tappet path or control path or height of the control pressure depends. At low pilot pressure, i.e. h .: As long as the Area of the pilot tappet 9 with the largest cross section within the Pilot channel 34 is located, this connection is very much throttled. at If the valve is opened further, the throttling effect becomes less than the throttling effect the compensating throttle 14 in the main piston. This results in a slow pressure drop in the pilot room 15 and thus begins a slow one 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. Therefore, the load lowers very slowly. The movement of the main piston 3 in the sense of opening the main valve seat means relative to the pilot piston and plunger 9 a movement in the sense of closing of the pilot valve 6/7, because the absolute position of the pilot valve 9 by the position of the control piston 20 is predetermined. Since the main piston 3 the pilot piston 8 follows in its movement, so the throttle cross section narrow again at the throttle point 36 in the pilot channel 34. As a result, a higher pressure builds up again in the pilot control chamber 15. This pressure build-up ensures that between the pilot piston 8 and the main piston 3 is set to an equilibrium state.

As soon as the control pressure is increased further, the area of the pilot tappet appears 9 with a decreasing cross-section further from the pilot channel 34 and the pilot valve seat 6. This becomes the pilot channel 34 opened further, d. h .: the throttling effect of the pilot tappet 9 continues reduced. An increasing volume flow flows from the pilot room 15 past the pilot ram 9, for. B. by the in the pilot tappet 9th arranged throttle grooves 10, in the return chamber 73. Here is the Throttle cross section in the pilot channel 34, z. B. by the throttle grooves 10, dimensioned such that with the movement of the pilot tappet 9 a uniform slow reduction of the throttling effect and consequently a steady one Pressure reduction in the pilot control room 15 occurs. This will make a progressive Opening behavior of the pilot piston 8 achieved by the size of the control pressure is clearly defined.

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

The design of the main piston 3 in connection with the valve seat 5 also has the advantage that the flow forces acting in the closing direction always counteracts a hydraulic opening force that in everyone Position is greater than the flow forces. With that, the impact possible pressure vibrations in port B on the main piston 3 avoided.

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

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

In lifting mode:

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

It should be noted that in the load holding brake valve according to the invention the compensating throttle 14 and the pre-throttle bore 41 also through Nozzles can be replaced so that a pressure-independent pressure reduction can be done.

A pressure relief valve for load securing can be placed in the load holding brake valve to get integrated. This is shown and described with reference to FIG. 4.

4 is with regard to the control chamber 2 and control bore 43 and the valve function identical to the load holding valve Fig. 2. Therefore, reference is made to the description there and just mentioned the differences.

In this embodiment, the pilot piston 8 and the main piston 3 advantageously clamped only with the spring 12, the supports the valve housing. The main piston 3 is essentially by hydraulic forces moved axially. The pilot piston 8 has this Execution of a guide shaft 37 which in the stepped bore 71 of the Main piston 3 is guided sealingly. This forms between the pilot valve seat 6 and guide shaft 37 concentrically to the control piston 8 Anteroom 40 to the pilot room 15. The antechamber 40 is a Pre-throttle 41 connected to the pilot room 15. The throttle cross section the pre-throttle 41 can be larger, equal to or smaller than the throttle cross-section the compensation choke 14 may be executed. This configuration of the 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 exhibit. The pre-throttling hole is particularly effective when open 41 that a higher closing force on the Pilot piston 8 acts. The higher closing force leads to the axial displacement of the throttle cross-section in the pilot channel (throttle point 36 in FIG. 3) is also reduced in size and thus the main piston 3 increasingly closes due to the overrun regulation. This system is special an advantage in an open circuit. Here is a control pressure predetermined on the control piston 19, which is independent of the Pump pressure and regardless of the inlet pressure, e.g. B. also adjustable is.

Due to the large area of the control piston 20 can in the control bore (Spring chamber) 43 as a control spring two connected in parallel preloaded springs 46 and 47 between the guide collar 20 and the Separator 17 can be clamped. If one spring breaks, the other is The spring is able to close the pilot piston in its starting position move. This is of particular importance in terms of security.

• daß der Druckbegrenzungskolben 55 einen Schaft besitzt, der die Lastkammer 53 durchdringt und als Ringraum ausbildet;
• daß die Lastkammer 53 einerseits durch den Kolben 55 mit Rückschlag-Ventilsitz 54 und an der anderen Seite durch einen am Schaft befestigten Endbund 62 begrenzt wird, und
• daß der Rückschlag-Ventilsitz 54 nur eine geringfügig größere hydraulische Wirkfläche hat als der am Schaft befestigten Endbund 62.

In the load holding 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 side (return space 73). The pressure-limiting piston 55 has only a very small area acting in the opening direction. This is achieved

• that the pressure limiting piston 55 has a shaft which penetrates the load chamber 53 and forms 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 fastened to the shaft, and
• that the check valve seat 54 has only a slightly larger hydraulic effective area than the end collar 62 attached to the shaft.

To design is in the valve housing 1 on the side of the control chamber a blind hole 50 is introduced at the end. The blind hole is 50 by means of overload bore 49 with the annular space (load chamber) 70 and over the Return bore 60 connected to the return chamber 73. In the blind hole 50 the plug 51 (socket) is screwed in. In the stopper 51, an inner bore 52 is made in the center, towards the blind hole is open and forms the check valve seat 54 with its end. The Check valve seat 54 is between the overload bore 49 and the Return bore 60. Inner bore 52 is via radial bores 53 and a recess 76 on the plug 51 with the overload chamber 49 connected. The overload chamber 49 and the return chamber 60 are between the bore 68 and the inner bore 52 of the valve housing Pressure relief valve 30 arranged. The spring-loaded pressure limiting piston 55 of the pressure relief valve 30 has a sealing surface 56, that under the biasing force of a compression spring 57, 66 on the check valve seat 54 rests and the radial bore 53 with respect to the return space 73 seals. The pressure limiting piston 55 has on both sides an end collar 62, 63. The piston shaft penetrates the radial bore 53 and has an end collar 62 at its end. This end collar is 62 in the inner bore 52 sealingly (seal 79), the end face 64 is slightly 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 is - With a tapered end part - in the front wall and guide hole 77 guided with seal 61 and protrudes into the bore 68. The inner bore 52, which is adjacent to the overload bore 49 and its end collar 62 are with the pressure of the return chamber 73 loaded. A relief channel is used for this 81, which is designed as a longitudinal bore in the axis of the piston and the the return chamber 73 through a radial branch channel 80 with the end space connects at end collar 62. The cross section of this end space as well as the End collar 62 is slightly smaller than the seat 54 of the check valve seat 54. The effective area, which acts in the opening direction at load pressure is this difference. The bore 68 is for pressure relief through the relief bore 69 with the control bore 43 (spring chamber) and the leak oil hole L. The one protruding into hole 68 thinner end collar 63 is in terms of its hydraulically effective cross section (Face 65) the same size as the above. Effective area in the opening direction, d. that is, the difference between the valve seat surface 54 and the cross section the inner bore 52 with end collar 62.

The piston 55 of the pressure relief valve 30 is connected in parallel with two Compression springs loaded in the closing direction; which is a compression spring 57 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 clamped. The plug is used to adjust the load securing pressure 51 screwed more or less deep into the blind hole.

About the function of the pressure relief valve 30:

The load pressure is in the inner bore 52 against the sealing surface 56 of the Valve seat 56 on. As soon as the set load securing pressure is reached, without first reducing the volume flow via the main piston 3, the pressure-limiting piston 55 becomes axial against the springs 57 and 66 postponed. The sealing surface 56 rises from the check valve seat 54, and the pressure relief valve 30 opens. The oil can now drain from the overload hole 49 via the opened valve seat 54 to the return bore 60 stream. As a result, the annular space 70 and the return chamber 73 bypassing the valve seat on the main piston 3 when the Load pressure exceeds a preset limit. The limit value (load securing pressure) is connected in parallel by the two Compression springs 66 and 57 specified.

The design of the pressure limiting piston 55 and its pressure relief has the consequence that the opening pressure in the inner bore 52nd acts on the valve seat 54, regardless of the return pressure and exclusively depends on the load pressure. This embodiment of a pressure relief valve is particularly important for the load securing function in the load holding brake valve suitable. Because in the usual circuits a downstream one If there is a pressure relief valve in the directional control valve, this occurs not to add up the set pressures.

Figures 5 to 10 show a possibility of hydraulic stroke limitation of a pilot operated valve. This hydraulic stroke limitation lets apply to all hydraulically pilot operated valves where a Control valve for actuating a valve piston is provided. The hydraulic Stroke limitation is illustrated on a load holding brake valve; how it is labeled in Figures 1 to 4. 5 is the circuit diagram Similar to the circuit diagram of FIG. 1. On the description of FIGS. 1 to 4 full reference is made. The pressure relief valve 30 after 3, 4 is not shown here. The load holding brake valve is added through a metering valve 84 in the control of the control piston 20 the control connection X.

A metering valve 84 is used for this control. The metering valve 84 is in Details are shown in FIG. 6 and will be described with reference to FIG. 6.

The metering valve 84 is located in the cover 22, which is the control chamber 21 limited. The cover 22 is on the valve housing 1 of the load holding brake valve flanged pressure-tight by means of seal 121. The dosing chamber with valve seat 109 of the metering valve 84 is relative to the control chamber 21 movably guided and positionable. For this purpose, e.g. B. the Valve seat 109 of metering valve 84 formed on a closing piston 119, which dic dosing valve chamber 102 compared to the control chamber 21 in other closures and which in the metering valve chamber 102 parallel to the pilot piston is sealingly guided and positionable. To do this 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 on hers End, which faces away from the load holding brake valve, with a Thread 105 provided. On their remaining length (link level 104) has a larger diameter. An adjustment spindle is in the thread 105 106 screwed in and pressure-tight with a counter sealing nut 113 braced. The adjusting spindle 106 forms with the longitudinal bore 104, 105 an annular space in the area of connection stage 104. In this connection stage 104 opens the tax line. In the control line X is a filter 116 and a nozzle 117 turned on. The annulus is on the Side facing the load holding brake valve by a guide collar 119 completed, which is firmly connected to the end of the adjusting spindle 106 and which in the guide step 103 of the longitudinal bore 102 by means of O-ring seals 120 is performed sealingly. The adjusting spindle 106 is penetrated centrally by a central channel 108. To the The end, which faces away from the load holding brake valve, is the Central channel 108 is closed pressure-tight by a plug 112. To the End of the central channel 108, which faces the load holding brake valve, opens the central channel 108 with a valve opening channel 107 in the Control chamber 21. This is located in front of the valve opening channel 107 Dosing valve with the closing element 110 and a stem 118. Das Closing element 110 is a ball here. The shaft 118 is supported on the one hand on the closing element 110 and is preferably with a control piston 20 firmly connected. The stem 118 also penetrates the valve opening channel 107 big game and protrudes into the control chamber 21, where it is at the front of the control piston 20, which the control chamber 21 to the other Side limited, is present. The central channel 108 forms with the diameter smaller valve opening channel 107 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 so by a spring 111 in the direction of the control piston 20 pressed that it is over the shaft 118 on the front side of the control piston 20 supports. In the depressurized state of the control chamber 21 is the Control piston 20 under the force of the springs 46, 47 on the cover 22 in which the metering valve is located. In this position the Shank 118 the closing element 110 so far from the valve seat 109 that a radial channel 114 has space which the bore 102 and the opening therein Control channel X via the connection stage 104 with the central channel 108 connects.

About the function:

If the control connection x is subjected to control pressure, then plants the control pressure in the connection stage 104 and the radial channel 114 to the central channel 108. Since the closing element 110 compared to the Walls of the central channel 108 has great play, the control pressure is too high both sides of the closing element 110. The oil flow then passes through Valve opening 107 into the control chamber 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 shaft 118 and closing element 110 the opening movement of the opening piston participate. The closing element 110, which is shown here as Ball is formed at the end of the central channel 108 and to the system to the valve seat 109 of the valve opening. This will open the valve 107 closed and the opening movement of the opening piston 20 completed.

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

When the control port is depressurized, on the other hand, the control piston moves 20 under the load of the springs 46, 47 except for his Stop, d. H. the lid 22 back.

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

a) Dosierbypass
Von dem Ringkanal 104, der mit dem Steuerdruck beaufschlagbar ist, zweigt ein Dosierbypasskanal 126 über eine Dämpfungsdüse 125 ab. Dieser Dosierbypasskanal 126 setzt sich in einem Stichkanal 127 fort, der in den Aufsteuerraum 21 einmündet. In dem Stichkanal 127 kann erforderlichenfalls eine weitere Dämpfungsdüse 128 angeordnet sein.
Zur Funktion:

Durch den Dosierbypasskanal 126 mit einer oder mehreren Dämpfungsdüsen 127 wird der Aufsteuerraum 21 auch dann mit dem Steuerdruck beaufschlagt, wenn das Schließelement 110 den Ventilsitz 109 schließt. Es erfolgt jedoch lediglich noch eine in gewünschtem Maße bedämpfte Aufsteuerung des Aufsteuerkolbens 20. Es ändert sich hierdurch also auch die Funktion des Dosierventils.

Bei Anwendung des Dosierbypass bewirkt das Dosierventil eine ungehindert schnelle Ansteuerung des Aufsteuerkolbens 20 in dem ersten Steuerbereich. Das Dosierventil bewirkt ein schnelles Ansprechen des Hauptventils, d. h.: des Lasthalte-Bremsventils im Absenkbetrieb. Dieser Schnellsteuerbereich ist beendet, wenn das Dosierventil den Zufluß des Steueröls durch Ventilöffnung 107 unterbindet (hydraulische Hubbegrenzung des Schnellsteuerbereichs). Nunmehr wird der Aufsteuerraum nur noch über den Bypass stark gedrosselt mit Steueröl beaufschlagt. Die Beschleunigungsabsenkung wird entsprechend verlangsamt. In diesem Zustand bleibt allein der Dosierbypass 127 wirksam, so daß das Lasthalte-Bremsventil feinfühlig betrieben werden kann. Ohne den Einsatz des Dosierventiles wären die Forderungen: großer Aufsteuerweg mit schneller Ansteuerung bei guter Dämpfung nur möglich, wenn in den Steuerkanal x die für die Dämpfung erforderlichen Düsen eingesetzt würden und zum schnellen Ansteuern der Aufsteuerbewegung ein sehr hoher Steuerdruck angewandt würde. Durch Verstellung der Verstellspindel kann das Verhältnis des Schnellsteuerbereiches zu dem Gesamtsteuerbereich eingestellt werden.

Andererseits wird durch den Einsatz des wenig drosselnden Dosierventils auch eine schnelle Rückbewegung des Aufsteuerkolbens 20 ermöglicht, da die beiden Dämpfungsdüsen 125, 128 im Dosierbypass 126 durch die Ventilöffnung 107 umgangen werden.

b) Tank-Bypass
Von dem Dosierbypass zweigt ein Tank-Bypasskanal 137 ab, der den Dosierbypass mit dem Tank-Kanal 138 verbindet. In dem Tank-Bypasskanal 137 ist eine Bypassdüse 132 sowie ein Bypassrückschlag-Ventil mit Kugel 133 und Feder 134 angeordnet. Das Rückschlagventil verhindert den Rückfluß im Dosierbypass 126 vom Leckölabschluß L über die Düse 132 zur Verbindungsbohrung.
Zur Funktion:

Der Druck in der Verbindungsbohrung 126 öffnet die Kugel 133 des Bypass-Rückschlagventils. Dadurch fließt ein Teil des Steueröls durch die Bypassdüse 132 und den Bypasskanal zum Tank. Es entsteht damit eine Strom- und Druckteilung im Dosierventil. Hierdurch werden Druckschwingungen gedämpft. Die Stärke der Dämpfung kann durch die Größe der Bypass-Düse 132 bestimmt werden.

c) Vorspannung des Steuerdrucks
Von dem Ringraum 104, der mit dem Steuerdruck belastet ist, zweigt ein Vorspann-Bypass 129, 131 ab. In diesem Vorspann-Bypass liegt ein Vorspannventil (Überdruckventil 130), das durch eine Schraube einstellbar ist. Das Vorspannventil weist - wie bekannt - ein federbelastetes Rückschlagventil auf, das durch den Druck in dem Ringkanal 104 geöffnet wird und eine Verbindung mit dem Aufsteuerraum 21 herstellt.
Zur Funktion:

Bei plötzlichem Anstieg des Steuerdrucks in dem Ringraum 104 und vor der Dämpfungsdüse 125 erfolgt eine Öffnung des Vorspannventils 130. Damit fließt das Steueröl schnell und direkt in den Aufsteuerraum 21. Es erfolgt eine schnelle Reaktion des Aufsteuerelementes im Sinne einer Öffnung des Lasthalte-Bremsventils für das Senken der Last.

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 can be used alone or in combination with two or three together with the metering valve:

a) Dosing bypass
A metering 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, control of the control piston 20 is only damped to the desired extent, thus also changing the function of the metering valve.

When using the metering bypass, the metering valve effects an unimpeded, rapid activation 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 inflow of the control oil 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: long control path with quick control with good damping 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 area to the total control area can be set.

On the other hand, the use of the less throttling metering valve also enables the control piston 20 to move back quickly, 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 connection 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 a screw. As is known, the preload valve has 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:

In the event of a sudden increase in the control pressure in the annular space 104 and in front of the damping nozzle 125, the preload valve 130 is opened. The control oil thus flows quickly and directly into the control space 21. The control element reacts quickly in the sense of an opening of the load holding brake valve for the Lowering the load.

While through the metering valve during normal operation of the load holding and Brake valve is already enabled by a quick control the preload valve used in combination with it further accelerated Activation bypassing the quick control area and the fine control area allows.

It should be noted that the metering valve by itself, but also in combination with one or more elements a, b and c also for other control tasks can be used which involve a control piston, by which a hydraulic flow is controlled, by a control pressure hydraulically controlled and adjusted, especially against the To be adjusted by means of a return spring.

LIST OF REFERENCE NUMBERS

1

valve housing

1A

Load-holding brake valve

2

control chamber

3

main piston

4

sealing surface

5

valve seat

6

Pilot valve seat

7

sealing surface

8th

pilot piston

9

Pilot tappet, piston skirt

9A

approach

10

throttle

11

spring plate

12

feather

12A

feather

13

Control chamber plug

14

interphase reactor

15

Spring chamber, pilot control chamber

16

Steering, thin part

17

divider

18

poetry

19

Taxation, big part

20

Opening piston

21

Opening chamber

22

Plug

23

center collar

24

Gating spring

25

lowering line

26

consumer

27

Return line

28

supply line

29

control line

30

Pressure relief valve

31

way valve

32

pump

33

tank

34

Pilot channel, seat bore

35

Undercut groove

36

restriction

37

guide shaft

38

guide bore

39

face

40

anteroom

41

Pre-throttle bore, pre-throttle

42

Dead end

43

Control bore, spring chamber

44

face

45

face

46

feather

47

feather

48

stop surface

49

Overload hole

50

Blind hole

51

Plug

52

internal bore

53

radial bore

54

Check-valve seat

55

Pressure limiting piston

56

sealing surface

57

compression spring

58

piston extension

59

hole bottom

60

Return bore

61

poetry

62

Dead end

63

Dead end

64

face

65

face

66

compression spring

67

Plug

68

drilling

69

relief well

70

annulus

71

stepped bore

72

step

73

Return Room

74

guide bore

75

face

76

recess

77

guide bore

78

recess

79

poetry

80

Stichkanal

81

relief channel

82

poetry

83

locknut

84

metering valve

101

cover

102

longitudinal bore

103

management level

104

connecting stage

105

thread

106

adjusting spindle

107

Valve opening channel

108

Central channel

109

valve seat

110

closing element

111

feather

112

Plug

113

Counter-sealing nut

114

radial channel

115

control terminal

116

filter

117

jet

118

shaft

119

closing piston

120

Seal (O-ring)

121

Seal (O-ring)

125

damping

126

Connecting hole, hole

127

Connection hole, damping channel

128

damping

129

Connection hole, preload channel

130

Pre-load valve

131

Connecting channel, prestressing channel

132

Bypass nozzle

133

Bypass check valve valve ball

134

Bypass check valve spring

135

Bypass channel, bore

136

Seal (O-ring)

137

bypass channel

138

Tank channel, connecting hole

141

Undercut groove

142

Area with the largest cross section

143

Area with decreasing throttling (throttling effect)

144

Partial length with decreasing cross section

145

Partial length with constant cross section

146

Area with the smallest cross section

147

header

Claims (18)

1. Hydraulically controllable load-holding brake valve, in particular for a double-acting consumer under an external load on one side on its load side, having the following features:

   a control chamber (2) is arranged in a valve housing (1);

   the control chamber consists of preferably aligned chamber portions in this sequence:

   pilot space (15);

   annular space (70) connected via the connection B to the lowering line (25) of the consumer (26);

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

   opening space (21) connected to a control duct (X);

   a valve seat (5) with a central passage by means of which the connecting bores A and B can be connected is fixedly arranged on the valve housing (1) in the control chamber (2) between the annular space and the return space;

   the valve seat is closed or opened by a main piston (3);

   the main piston (3) is in the form of a step piston and comprises:

   a thin piston collar which, together with the cylindrical wall of the control chamber (2), forms the annular space (70), a sealing surface (4) on the thin piston collar directed towards the valve seat and cooperating with the valve seat (5);

   a thick piston collar which is guided in a sealing manner along the wall of the control chamber between the annular space and the pilot space and separates the two from one another;

   the main piston is axially displaceable in the control chamber (2) by pressurisation of the return chamber (73) or the annular space (70) in the direction of lifting away from the valve seat (5) and by pressurisation of the pilot space (15) in the direction of closure of the valve seat;

   the pilot space (15) can be connected via a compensating throttle (14) to the annular space (70) and the connection B and via a pilot duct (34) with the pilot valve seat (6) in the main piston (3) to the return space (73) and the connection A;

   the pilot duct with the pilot valve seat (6) can be closed by a closing element guided concentrically with the pilot duct (34), the pilot piston (8), by means of its sealing surface (7) by pressurisation in the pilot space (15) and, preferably, the force of a closing spring (12) and can be opened in the opposite direction by a pilot tappet (9), the said pilot tappet (9) being guided with play in the pilot duct (34) and projecting into the return space (73);

   an opening piston (20) is axially guided in the opening space (21) and is displaceable in the direction of the return space (73) by pressurisation of the opening space (21) and in the opposite direction by an opening spring (24);

   the opening piston (20) has an opening shank (19) directed towards the pilot tappet (9) and aligned coaxially therewith which projects via one end, the opening end (16), into the control chamber (2) and acts on the pilot tappet (9) and on the pilot piston (8) in the direction of opening in the event of axial displacement of the opening piston (20) against the force of the opening spring (24),

   characterised in that

   the pilot tappet (9) has at least the following longitudinal regions over its length and departing from the bearing area (7) of the pilot piston (8):

   first a region (142) with a maximum cross section which is guided with minimum play relative to the pilot duct (34);

   then a throttling region (143) adjacent thereto which forms a throttle gap relative to the pilot duct (34) over its length with its cross section, said throttle gap departing from the throttle gap with the maximum cross section and then increasing continuously, preferably gradually, at least over part (144) of the length of the throttling region (143);

   then a region (146) with a minimum cross section;

   the pilot tappet (9) is preferably rigidly connected to the pilot piston (8),

   the active area (45) of the opening piston (20) acted upon by the control pressure is in a ratio of more than 50:1, preferably more than 100:1 to the active area of the pilot valve seat (6) and the ratio of the end face (45) of the opening piston (20) to the end face (44) or active area (74) on the opening end (16) is preferably more than 30:1, in particular more than 60:1,

   the throttle cross section (36) formed by the pilot tappet (9) and the pilot duct (34) is smaller than the opening cross section formed between the pilot valve seat (6) and the sealing surface (7) of the pilot piston (8) in all of the opening positions of the pilot piston (8), and

   the maximum throttle cross section formed by the pilot piston (8) and the pilot duct (34) is larger than the flow cross section of the compensating throttle (14).
2. Valve according to claim 1, characterised in that the pilot tappet (9) is circular cylindrical in the throttling region closest to the pilot piston and substantially has the maximum cross section and little play relative to the pilot duct (34), the pilot tappet (9) has at least one axially directed throttle groove (10) on its jacket in the following throttling region (143) with a decreasing throttling effect, the depth and/or width of which is substantially zero adjacent to the region of the maximum cross section and increases continuously over part (144) of the length of the throttling region (143) and, preferably, then continues in a constant manner over a further part (145) of the length, and, preferably, the bottoms of the throttle grooves at the other end of the throttling region (143) end substantially at the minimum cross section of the pilot tappet (9).
3. Valve according to claim 1 Or claim 2, characterised in that the main piston (3) is provided at its end directed towards the pilot space (15) with a central guide bore (38) from the bottom of which the pilot duct (34) departs;
   the pilot piston (8) is provided at its end directed towards the pilot space with a
   guide shank (37) which is guided in a sealing manner in a guide bore (38) in the main piston (3) and has an end face (39) larger than the active area of the pilot valve seat (6);
   the part of the guide bore (38) situated between the pilot valve seat (6) and the guide shank (37) is connected via a pre-throttle bore (41) to the pilot space (15).
4. Valve according to one of claims 1 to 3, characterised in that the annular space (70) including the connection B and the lowering line (25) and the return chamber (60) with the return space (73) including the connection A, the return line (27) and the tank are connected via a chamber (49) and a return chamber (60) and a spring-loaded pressure-limiting piston (55) of a pressure-limiting valve (30) arranged therebetween.
5. Valve according to claim 4, characterised in that the chamber (49) and the return chamber (60) are situated between two end chambers of the pressure-limiting valve (30), the spring-loaded pressure-limiting piston (55) of the pressure-limiting valve (30) has a sealing surface (56) and, at either end, a piston shank with a guide end (62) and a guide end (63), the sealing surface (56) bearing against the valve seat (54) under the prestressing force of a compression spring (57) and each guide end (62, 63) being guided in a sealing manner in one of the end chambers of the valve housing (1), the said end chamber adjacent to the overload bore (49) and its guide end (62) are loaded by the pressure of the return chamber (73) via the longitudinal bore (81) and the transverse bore (80) and have a slightly smaller cross section (64) than the bearing area of the valve seat, and the end chamber with the guide end (63) is relieved from pressure and its hydraulically active cross section (65) is identical to the difference between the valve bearing area (54) and the cross section (64) of the end chamber (147) with the guide end (62).
6. Valve according to claim 4 or claim 5, characterised in that the spring-loaded pressure-limiting piston (55) of the pressure-limiting valve (30) is loaded by two compression springs connected in parallel, one compression spring (57) being fixed relative to the pressure-limiting piston (55) in the return chamber (60) and the other compression spring being fixed relative to the piston shank with the guide end (63) in the end chamber relieved from pressure.
7. Valve according to one of claims 1 to 6, characterised in that the opening chamber (21) is connected to the control duct (X) via a metering valve (84) by means of which the opening piston (20) is acted upon by a quantity of control oil limited to a predetermined stroke of the opening piston (20).
8. Valve according to claim 7, characterised in that the metering chamber (102) of the metering valve (84) is connected to the control connection (115) and comprises:

   a valve opening (107) with a valve seat (109) through which the control oil reaches the opening space (21), a closing element (110) which is supported by means of a shank (118) on the opening piston (20) in such a manner that the closing element (110) is movable in the metering chamber (102) between the valve seat (109) and an opening position synchronously with the opening piston (20), and that it closes the valve seat (109) given a predetermined stroke of the opening piston (20).
9. Valve according to claim 8, characterised in that the valve opening (107) in the opening space (21) opens on to the side directed away from the opening spring (147) and is surrounded by an annular closing surface (109) situated parallel to the end face of the opening piston acted upon by pressure, the shank (118) passes through the valve opening (107) with large play, and the closing element (110) is pressed against the valve seat (109) by a spring (111) bringing the shank to bear against the end face of the opening piston (20) acted upon by pressure and after executing the predetermined stroke.
10. Valve according to claim 9, characterised in that the valve seat (109) with the valve opening (107) of the metering valve (84) is movably guided and can be positioned relative to the opening chamber (21), in particular the valve opening (107) of the metering valve (84) is formed on a closing piston (119) which closes the metering valve chamber (102) relative to the opening space (21) and which is guided in a sealing manner and can be positioned in the metering valve chamber (102) parallel to the opening piston (20).
11. Valve according to claim 10, characterised in that the closing piston (119) can be positioned in such a manner that it bears against the opening piston (20) and displaces and positions the opening piston (20) in the direction of unblocking the pilot closing element (8).
12. Valve according to claim 10 or claim 11, characterised in that the closing piston (119) is mounted on the free end of an adjusting spindle (106), the adjusting spindle (106) has a central duct (108) aligned with the valve opening (107) and closed at the free end of the adjusting spindle (106) by the plug (112), the closing element (110) is guided in the central duct (108), control pressure acts upon the central duct (108) on either side of the closing element and the adjusting spindle (106) can be screwed into or unscrewed from a threaded bore (105) parallel to the movement of the opening piston (20).
13. Valve according to claim 12, characterised in that, in one end position of the adjusting spindle (106), the closing piston (119) projects into the opening chamber (21), bears against the opening piston and displaces the opening piston (20) in the direction of unblocking the pilot closing element (8) (Fig. 8) and, in the other end position, the distance between the valve seat (109) and the end face of the opening piston (20) situated in the rest position is shorter than the shank (118).
14. Valve according to claim 12 or claim 13, characterised in that the control pressure acts upon the central duct (108) on either side of the closing element (110), the control pressure duct (114) opening into the central duct (108) directly in front of the closing surface of the closing piston and the closing element being guided with play in the central duct.
15. Valve according to claims 7 to 14, characterised in that the shank (118) is rigidly connected to the closing element (110) or is separate from the closing element (110), and that the shank (118) is rigidly connected to the opening piston (20) or is separate from the opening piston (20).
16. Valve according to one of claims 7 to 15, characterised in that the metering valve (84) is by-passed by a throttling duct (127) which displays greater throttling (125, 128) of the control oil flow after the closure of the seat (109) by the closing element (110).
17. Valve according to one of claims 7 to 16, characterised in that the metering valve is by-passed by a pressurising duct (129, 131) with a pressurising valve (130) disposed therein, by means of which the maximum pressure difference between the pressurising duct (129) and the remaining control duct (21) is predetermined.
18. Valve according to one of claims 7 to 17, characterised in that the metering valve is by-passed by a relief duct, such as by-pass ducts (135, 137), which connects the control duct to the tank via a by-pass nozzle (132) and a non-return valve (133).

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