WO2009071264A1

WO2009071264A1 - Hydraulic system having adjustable rapid-lowering valve

Info

Publication number: WO2009071264A1
Application number: PCT/EP2008/010200
Authority: WO
Inventors: Seppo Tikkanen
Original assignee: Robert Bosch Gmbh
Priority date: 2007-12-04
Filing date: 2008-12-02
Publication date: 2009-06-11
Also published as: DE102008034301B4; DE102008034301A1

Abstract

The invention relates to a hydraulic system, comprising a hydraulic actuator (1), comprising at least one first chamber (a) and a second chamber (b) having a connection each for connecting to the hydraulic system, at least one adjustable hydraulic pump (2), comprising at least one first connection and a second connection, wherein the first connection is connected to the first chamber (a) of the hydraulic actuator and the second connection to the second chamber (b) of the hydraulic actuator (1). The first chamber (a) is connected to the second chamber (b) via at least one rapid-lowering valve (7) having a controllable flow cross-section.

Description

Hydraulic system with an adjustable quick-release valve

The invention relates to a hydraulic system with a hydraulic actuator and at least one adjustable hydraulic pump.

In known pump-controlled hydraulic systems of a shovel loader or the like, the flow rates and thus e.g. the lifting and lowering speed are specified by the pump (s) in the system. During a normal work operation of a machine, in the case of a bucket, a gravitational force always acts on the boom of the bucket loader, that is, in the hydraulic system, for example, a load pressure always acts on a bottom side of the boom cylinder. In this case, the lifting speed as well as the lowering speed is controlled by the flow rate on the bottom side of the cylinder. Generally, in the operation of a machine controlled by such a hydraulic system, the lowering speed should be 30% to 70% higher than the lifting speed to allow efficient operation of the machine. However, this requirement for the asymmetrical operation generates a nominal size problem in the pump-controlled hydraulic system. If the pump is designed according to the lowering speed, the resulting pump has a very large nominal size. With this pump, however, the maximum lifting speed can not be achieved when lifting at full load, because the available power of the engine is not sufficient. Consequently, during the load stroke, the swing angle of the pump would only be about 50% and thus the efficiency of the pump would not be in the optimum range.

This problem leads to the desire to develop a hydraulic system for controlling lifting cylinders or comparable hydraulic actuators, which has a equal to the lifting speed faster lowering speed of the lifting cylinder allows or reduces the nominal size of the pump at the same lowering speed. This achieves better pump efficiency when lifting as well as lowering a lifting cylinder.

From US 6,804,957 B2 a hydraulic system is known that allows a quick lowering function of a hydraulic cylinder. The system includes a bypass valve located between the bottom of the hydraulic cylinder and the pump and connected directly to the tank. The valve can be opened in the lowering operation of the cylinder and directs part of the volume flow directly into the tank. The derived oil volume is therefore missing in the system and an additional pump is needed as a storage loading pump, which promotes the oil back into the circulation or into the storage tank, which requires additional energy. In order to enable energy-efficient and efficient work, however, a hydraulic system would be desirable that does not require an additional pump and in which the nominal size of the pump by the required operating speed against the load, for example when lifting a bucket or the pressing of garbage in refuse collection vehicles , is determined.

The aim of the present invention is therefore to achieve a higher speed in the load direction in a pump-controlled hydraulic working circle at the speed against the load direction predetermined nominal size of a hydraulic pump or vice versa to reduce the nominal size of the pump at the same lowering speed, without using an additional pump and thus to increase energy consumption in the system.

The object is achieved by the hydraulic system according to the invention according to claim 1 and the method according to claim 12. In a first aspect, the invention relates to a hydraulic system having at least one hydraulic actuator, for example for lifting loads, comprising at least two chambers for moving the actuator in or against the load direction. The chambers each have a connection for connection to the hydraulic system. The hydraulic system comprises at least one adjustable hydraulic pump with at least two ports, one port being connected to a first chamber of the hydraulic actuator and the second port being connected to a second chamber of the hydraulic actuator. Furthermore, the hydraulic system has at least one fast-flow valve with a controllable flow cross-section, via which both chambers can be connected.

With such a hydraulic system, it is possible to reciprocate pressure means between the chambers of the hydraulic actuator by means of the pump, thereby enabling actuation, for example lifting and lowering. Since the actuator can be installed in a variety of ways in a machine, for example, it would also be conceivable to arrange the actuator horizontally, whereby an extension and retraction of the actuator is made possible by the hydraulic pump of the system described above. In the following, for the sake of simplicity, we will talk about lifting and lowering a load.

Since the quick-lowering valve is arranged parallel to the pump and connects the two chambers of the actuator with each other, the quick-lowering valve is used as a by-pass circuit for the pump to increase the total volume flow and thus the operating speed. It is therefore possible to convey a portion of the pressure medium not by the pump, but by an existing pressure difference due to a load from the first chamber of the actuator directly into the second chamber of the actuator. Since in this case no longer the entire pressure medium must be pumped by the pump, but a part of the Pressure medium is pumped bypassing the pump through the quick-flow valve, it is possible to achieve an overall larger volume flow of the pressure medium. Accordingly, the quick-flow valve can be used to achieve, for example, a higher lowering speed when lowering the actuator. The increase in the operating speed is generally possible when the actuator is actuated in the load direction.

An advantage of the above-described design of the system is that, unlike conventional systems, there are no additional lines between the quick-release valve and a tank of the hydraulic system. In addition, can be dispensed with a feed pump that promotes pressure fluid from the tank back into the lines. Consequently, the energy consumption of the system can be reduced compared to other systems.

Preferably, the quick-flow valve is opened only when the pump is fully swung out. This already creates a pressure difference between the chambers of the actuator before the opening of the quick-release valve. Preferably, the flow cross-section of the quick-release valve is dimensioned so that the setting of a pressure drop over the quick-flow valve is achieved. It is thus possible to pass a part of the volume flow through the quick-flow valve.

The flow cross section of the quick-release valve is preferably freely adjustable in a value range from 0% to 100%. It is therefore possible to influence the volume of oil delivered by the quick-flow valve or the volume flow with the aid of the adjustment of the flow cross-section. The lowering speed or operating speed in the load direction of the actuator can thus be adapted individually to the wishes of a user. With the system described above, different speeds of the actuator can be achieved, depending on whether and how far the fast-flow valve is opened. The pump delivers the same maximum oil volume when lifting or lowering the actuator, regardless of whether the quick-lowering valve is open or not. It is therefore possible to operate the pump in both modes of the actuator, in the lifting as well as in the lowering operation close to their optimum efficiency.

Preferably, in the stroke operation of the actuator, the quick-flow valve is closed, whereby a by-pass circuit of the actuator or the pump is prevented. In the lifting operation, therefore, the entire pressure medium to be pumped is conveyed only with the help of the pump in one of the chambers.

The pump of the hydraulic system described above is preferably a dual-circuit pump and has a nominal size, which is designed according to the required speed against the load, that is, for example, the stroke speed of the actuator. As a result, the pump is smaller and cheaper than when dimensioning the pump according to the required speed in the load direction, ie the lowering speed, which is generally above the required lifting speed.

Also possible is the integration of a memory for pressure medium in the system described above. This allows the supply of pressurized fluid. The pressure medium is stored there when lowering from a load increasing the stored pressure energy.

In a further preferred embodiment, the hydraulic system according to the invention also has a flow compensating valve, via which the first or the second chamber of the actuator can be connected to a tank volume, preferably via a pressure limiting valve. The quantity compensation valve is designed such that an excess of pressure medium in one of the chambers of the actuator or in one of the lines connecting the chambers of the actuator to the pump is prevented during operation of the quick-release valve. This prevents pressure equalization between the two chambers of the actuator and thus ensures the function of the quick-release valve.

In another aspect, the present invention describes a method of actuating a hydraulic actuator by means of a hydraulic system. In the method according to the invention, on the one hand, a volume flow is generated from a first chamber of the hydraulic actuator into a second chamber of the hydraulic actuator by a pump. In addition, the volume flow is increased to further increase the speed by opening a fast scavenge valve connecting the first to the second chamber. The erfindungemäße method has the advantage that no longer the entire required for lowering a load volume flow from the first chamber must be funded in the second chamber by the pump. Rather, a portion of the total volume flow is conveyed via the quick-flow valve directly from the first chamber into the second chamber. Accordingly, the pump can be made smaller.

According to an advantageous embodiment, the movement of the actuator in the load direction is divided into two areas. First, to increase the actuation speed, the pump is first swung out in the direction of its maximum delivery volume. After reaching the maximum delivery volume of the pump, to further increase the actuation speed of the hydraulic actuator, an adjustable quick-release valve is moved toward its open position, thus providing an increasingly open connection between the first chamber and the second chamber of the hydraulic actuator. The total volume flow from the first chamber into the second chamber is thus exceeded by the maximum delivery rate. volume of the pump increases. This makes it possible to achieve an increased speed in the direction of the load compared to the actuation speed of the actuator against the load direction, that is to say, for example, when lifting a load.

In this case, the user preferably selects from two control ranges, wherein the controllable fast-flow control valve can be opened in at least one of the two control ranges. The user thus has the choice between the first control range, in which the conveying speed of the pump can be adjusted continuously, and a second control range, in that the fast-flow control valve can be actuated at the maximum delivery rate of the pump.

In this way, it is ensured that the pump already delivers maximum when opening the quick-release valve and thus has already set a pressure gradient between the two chambers of the actuator. In addition, a user thereby has the choice of having the actuator lowered only by means of the pump or additionally adding the quick-flow valve for a faster lowering of the actuator. A certain overlap area, in which the delivery volume of the pump increases at the same time and the fast-flow valve is already opened, is also conceivable.

With opposite movement of the actuator, the user can also adjust the speed of the pump continuously. The quick-lowering valve can not be opened during this movement.

Furthermore, the quick-lowering valve is automatically closed in the lowering operation, if the load pressure exceeds a certain limit. It can thus be prevented that, for example, a machine boom under load during quick lowering operation represents a danger to the user or bystanders. According to the invention, an overpressure in one of the two chambers of the actuator can be prevented by a quantity compensation valve integrated in the hydraulic system. This compensates for an excessive pressure medium flow generated by the hydraulic pump. The resulting difference quantity is discharged via the flow compensating valve into the tank.

A preferred embodiment of the hydrostatic system according to the invention is illustrated in the drawings and will be explained in more detail in the following description. Show it:

1 is a schematic representation of the hydraulic system according to the invention according to a preferred embodiment,

2 is a schematic representation of the hydraulic quick-lowering valve according to the invention,

Fig. 3 is a schematic representation of the hydraulic quick-release valve according to the invention with float valve

4 shows a schematic representation of the hydraulic system according to the invention with a quantity compensation valve and discharge of the pressure medium via a pressure-limiting valve of the feed device, and

Fig. 5 is a schematic representation of the hydraulic system according to the invention with a flow compensating valve and separate pressure relief valve.

1 has an actuator 1 with two chambers a and b. Further the actuator consists of a piston rod Ia and a piston Ib. The separation of the two chambers a, b is carried out by the piston Ib. The second chamber b is arranged on a piston rod side of the actuator 1 and the first chamber a is arranged on the bottom side of the piston 1 b of the actuator 1. By introducing pressure medium into the first chamber a, the actuator 1 can thus be lifted. The introduction of pressure medium in the second chamber b allows a lowering operation of the actuator.

In the embodiments of the invention, a lifting and lowering of a load is assumed. In general, however, it is the case that the hydraulic actuator can be actuated counter to a load direction, which corresponds to the lifting, or can be actuated in the load direction, which is the case when lowering. This can be the case, for example, in a waste compactor, where work is performed against the load direction when compacting the refuse. It is assumed that for actuation against the load direction pressure medium of the first chamber a is supplied.

It should be noted that instead of the illustrated actuator 1, another hydraulic actuator 1 can be used. Thus, it is conceivable, for example, to use a hydraulic actuator 1 with a continuous piston rod Ia in order to obtain on both sides of the piston Ib of the actuator 1 equal pressure surfaces for the application of a pressure medium.

Furthermore, the hydraulic system according to the invention in Fig. 1, a pump 2, which is driven by a motor 3, preferably a diesel engine. The pump 2 is designed to be adjustable in terms of its stroke volume and preferably a hydraulic axial piston machine in swash plate construction, which is pivotable from a neutral position in two directions. The stroke volume is adjusted by tilting the swashplate. In the illustrated example, the pump 2 is a dual-circuit pump.

The dual-circuit pump has two individual pumps 2a, 2b. The first individual pump 2 a is connected via a storage line 15 to a hydraulic accumulator 4. With its second connection, the first individual pump 2a is connected to a first actuating pressure line 16, which connects the first individual pump 2a to the first chamber a. The second individual pump 2b also has two connections. A first connection is connected via a connecting line, which opens out in the first control pressure line 16 with the first chamber a. By contrast, the second connection of the second individual pump 2b is connected to the second chamber b via a second control pressure line 17.

The first control pressure line 16 and the second control pressure line 17 are connected to each other via a first connecting line 18. In the first connecting line 18, a quick-flow valve 7 is arranged so that a throttled connection between the first control pressure line 16 and the second control pressure line 17 is produced in dependence on the setting of the quick-lowering valve 7.

Further, parallel to the connecting line 18, a second connecting line 19 is formed, which connects the first actuating pressure line 16 with the second actuating pressure line 17. In the connecting line 19, two oppositely arranged check valves Vl and V2 are provided. Between these two check valves Vl, V2, which respectively open in the direction of the first actuating pressure line 16 to the second actuating pressure line 17, opens a feed pressure line 20 in the second connecting line 19.

The remote from the second connecting line 19 end of the feed pressure line 20 is connected to a feed pump 6. The feed pump 6 sucks from a tank volume 5 Pressure medium via a suction line 21 at. The feed pressure line 20 can be connected via a pressure limiting valve 11 to the tank volume.

The feed pump 6 is also driven by a drive shaft connected to the pump 2 motor 3. The valves V 1 and V 2 are check valves which permit a flow only in the conveying direction of the pump 6 and thus make it possible to supply pressure medium in the event of pressure medium leakage.

As already stated above, the first control pressure line 16 is connected to the connecting lines 18 and 19. The second actuating pressure line B is connected to the respective other ends of the connecting lines 18 and 19. Thus, the hydraulic actuator 1 and arranged in the first connection line fast-lowering valve 7 are connected in parallel.

In the following, the operation of the hydraulic system with the quick-lowering valve 7 will now be explained in more detail.

During lifting operation of the actuator 1, the pump 2 conveys pressure medium into the first actuating pressure line 16. The rapid-flow control valve 7 is closed in the lifting mode. Thus, there is no promotion of the pressure medium in the direction of the second chamber b. The pressure medium passes into the first chamber a of the actuator 1. At the same time b is conveyed from the second chamber to the pump 2. The result is a lifting or extension of the piston Ib of the actuator. 1

In addition, 6 pressure fluid can be provided from the tank with the help of the feed pump to compensate for pressure medium leakage.

In the lowering operation of the actuator 1, the pump 2 delivers pressure medium from the first chamber a of the actuator 1 and into the second chamber b. Because of the piston rod Ia are the Volume flows from / into the first chamber a and in / out of the second chamber b different. The resulting difference volume is promoted by increasing the potential energy in the hydraulic accumulator 4. The result is a lowering or retraction of the actuator 1. Since it can come between the removal of pressure fluid from the hydraulic accumulator 4 and the return to a volume difference, as can be seen from the example below, is a way to charge the hydraulic accumulator 4 provided. This could be done for example by the feed pump 6, which would then be connected via a check valve to the hydraulic accumulator 4.

In order to increase the lowering speed, the quick-lowering valve 7 is provided. The quick-flow valve 7 creates a throttled connection between the first control pressure line 16 and the second control pressure line 17. In addition to the volumetric flow, which is conveyed from the first chamber a into the second chamber b due to the pump pivot angle, it can thus be achieved with at least partially opened rapid-flow valve 7 pressure medium from the first chamber a flow bypassing the hydraulic pump 2 into the second chamber b. The total volume flow from the first chamber a into the second chamber b thus increases, as a result of which the lowering speed is likewise increased.

The quick-lowering valve is, for example, an adjustable 2/2-way valve with proportional control between two end positions. In this case, a control signal proportional to a control signal is generated, which acts against the fast lowering valve 7 against a fast-lowering valve 7 acting in its closed position return spring.

Preferably, a regulation of the quick-release valve 7 is designed such that the opening of the quick-release valve 7 by a user is possible only at maximum delivery rate of the pump 2. Thus, a user can do that Pump volume of the pump 2 stepless select up to the maximum delivery volume. In addition, the user can open and close the quick-flow valve 7 with maximum delivery of pressure medium by the pump 2 if necessary. Accordingly, an acceleration of the lowering operation is achieved, since in addition to the pump 2 pressure fluid from the chamber a is conveyed through the quick-flow valve 7.

Since the pump 2 already provides its maximum delivery capacity before the fast-poppet valve 7 is opened, a pressure gradient arises between the first chamber a and the second chamber b of the actuator 1. This pressure gradient allows the quick-poppet valve 7 to work correctly. Because of the pressure difference in the chambers a and b There is also a pressure gradient in the control pressure lines 16 and 17 and thus also in the line 18. When opening the quick-lowering valve 7 thus flows a volume flow Q from the first control pressure line 16 into the second control pressure line 12. This volume flow Q is higher, the greater the pressure gradient between the lines 16 and 17. Preferably, therefore, a small valve cross-section of the quick-release valve 7 is to be selected, which allows a significant pressure drop and thus a high volume flow Q above the quick-flow valve 7.

The quick-flow valve 7 is equipped with a variable flow cross-section. A user can thus regulate the opening of the valve between 0 and 100% and thus influence the lowering speed of the actuator 1. The control of the quick-release valve 7 is preferably carried out electromagnetically. In addition, other control possibilities for the quick-flow valve 7 may be provided, as shown by way of example in FIG.

For safety reasons, a limitation of the lowering speed can be provided, which prevents the opening of the quick-release valve 7 at a certain load pressure or brings an already opened fast-flow valve 7 back to its starting position, in which the first Connecting line 18 is interrupted. In this case, for example, an integrated pressure sensor can measure the load pressure, that is to say the pressure in the first chamber a of the actuator 1, and pass this on to a control unit of the hydraulic system. If the load pressure exceeds a predetermined limit value, the quick-release valve 7 closes automatically. Preferably, this limit value corresponds to 1.2 times the pressure level of the actuator operation without load.

FIG. 2 shows a preferred connection of the quick-release valve 7, which is located in the first connection line 18. It is in the illustrated embodiment, a 2/2-proportional valve. In the initial position, which is set by a return spring 7b, the connection between the control pressure lines 16 and 17 is blocked. When controlling the quick-lowering valve 7 by the control valve 8, a throttled flow through the quick-flow valve 7 is made possible. In this case, pressure medium passes from the first control pressure line 16 to the second control pressure line 17. The flow in the reverse direction is blocked in the reverse direction even when the quick-flow control valve 7 is open due to an integrated non-return valve function. The pressure line 7a shown enables the above-mentioned safety function on hydrostatic way. When the load pressure is increased, the hydraulic force acting in addition to the spring force of the return spring 7b increases in the closing direction. Thus, the opening force required for opening also increases, and at constant opening force, the quick-lowering valve 7 is automatically brought to a closed position.

For pilot control, a control valve 8 is connected to the quick-flow valve 7 through the pressure line 8a. In this case, a pressure surface of the quick-release valve 7 is acted upon by the pressure line 8a with a hydrostatic control force against the force of the return spring 7b. A realization of different flow rates due to different switching positions of the rapid lowering valve 7 is thereby possible by adjusting the hydraulic control force.

The control valve 8 is preferably a 3/2 proportional valve with electromagnetic control. With the help of an additional spring 8c, a starting position of the valve is fixed. In the initial position, which forms a first end position of the valve, the control valve 8 according to FIG. 2 is therefore connected to the tank 5. In this switching position thus no pressure fluid from the control valve 8 through the line 8 a is promoted to the quick-lowering valve 7 and the pressure surface of the quick-release valve 7 is relieved. When the control valve 8 is actuated with the aid of the electromagnet 8d connected thereto, the pressure line 8a is increasingly connected to a pressure medium source 8b as a function of the height of the control signal. With an adjustment of the control valve 8 in the direction of this second end position can therefore be promoted pressure medium from the pressure medium source 8b to the pressure line 8a, whereby a targeted control of the quick-release valve 7 takes place in the opening direction. Due to the variable adjustability of the valve, the height of the pressure surface of the Schnellsenkventils 7 acting pressure is set.

The promotion of pressure medium from the source 8b may be realized for example by a feed pump. Since the control valve 8 is a proportional valve, different switching positions and thus also different flow rates can be realized by the control valve 8 with the aid of the control by the electromagnet 8d.

FIG. 3 shows an embodiment according to the invention of the quick-lowering valve 7 combined with an additional floating-position valve 9. The floating-position valve 9 is arranged in the second connecting line 19. In the illustrated float valve 9 is 2/2-way valve. In the starting position the swimming pool is Positioning valve 9 locked. This starting position of the valve is determined by the spring 9a. In addition, a second pressure line 9b is provided, which connects the floating position valve 9 with the first connection line 18 on the side of the quick-release valve 7 directed to the first chamber a. The spring force of the spring 9a and the force of the pressurization on the float valve 9 by the pressure in the line 9b add up. The electromagnetic control of the float valve 9 is thereby overridden hydraulically at elevated load pressure. The float position valve 9 thus goes into its closed position for safety reasons when the load pressure rises above a value that can be determined by the magnitude of the electromagnetic force.

The electromagnetic force is generated by an electromagnet 9c as a counterforce to the spring 9a and to pressurize through the line 9b. Upon application of the electromagnetic force and an uncritical low load pressure, an opening of the float valve 9 is effected. Due to the open switching position of the float valve 9, a floating position of the hydraulic system is realized with simultaneous opening of the quick-release valve 7. In this case, the first connecting line 18 of the hydraulic system is connected to the tank 5 through a throttle circuit 10 on the side lying to the second chamber b with respect to the quick-flow valve 7. The flow of pressure medium to the tank 5 is thus limited. The throttle circuit 10 is designed with a check valve so that, if appropriate, pressure medium can be sucked from the tank 5, so that, for example, cavitation is prevented in the hydraulic system.

FIG. 4 shows an embodiment according to the invention of the hydraulic system according to FIG. 1 with a quantity compensation valve 23. This is connected via a first connecting line 22a to the actuating pressure line 16 and thus to the first chamber a of the actuator 1. Via a second connecting line 22b, the quantity compensating valve 23 is connected to the actuating pressure line 17 and thus to the second chamber b of the actuator 1. The quantity compensation valve 23 is also connected to the feed system and thus connected via the pressure relief valve 11 to the tank 5.

The quantity compensation valve 23 according to the invention is preferably a 3/3-way valve which is actuated by occurring pressure forces or pressure differences. Depending on the pressure conditions occurring on the piston and the bottom side of the actuator 1, either the first chamber a or the second chamber b is connected to the tank 5 via the pressure relief valve. At pressure equilibrium, the quantity compensation valve 23 is in its starting position.

The quantity compensation valve 23 shown in FIG. 4 is a 3/3-way valve. In the starting position, the flow compensating valve 23 is blocked and the connecting lines 22a and 22b are separated from each other and from the tank 5. Two pressure measuring lines 25a, 25b are connected to the control pressure lines 16 and 17, whereby in the case of an overpressure in one of the two control pressure lines 16,17 the valve 23 is activated. This is done in such a way that the actuating pressure line 16 is connected to the tank 5 via the pressure limiting valve 11 by an overpressure in the actuating pressure line 17. At an overpressure in the control pressure line 16, however, the control pressure line 17 is connected to the tank 5 via the pressure relief valve 11. Accordingly, oil volume can be removed from the control pressure line 16 or from the control pressure line 17 to the tank 5 by actuation of the quantity compensation valve 23.

The quantity compensation valve 23 is preferably activated when the ratio of the current fast lowering speed to the maximum speed of the actuator 1 (in operation without fast-flow valve) exceeds a factor K. The factor K is calculated as follows:

- ¹

2 (A_st / A_bo-1) where A_st is the area of the rod side of the actuator 1 and A_bo is the area of the bottom side of the actuator 1.

The operation of the quantity compensating valve 23 will be explained below.

If the ratio of the quick lowering speed to the maximum speed of the actuator 1 is less than the previously defined factor K, less oil volume is conveyed from the rod side and thus from the first chamber a to the second chamber b than is needed in the second chamber b. This means that the delivered volume of the quick-release valve 7 and the pump 2 is not sufficient to cover the needs of the rod side (chamber b) of the actuator 1. Therefore, in this case, additional oil volume is conveyed to the chamber b by means of the feed pump 6. The flow compensating valve 23 is not activated in this case.

If the ratio of the fast lowering speed to the maximum speed of the actuator 1 is more than the previously defined factor K, more oil volume is conveyed from the rod side and thus from the first chamber a to the second chamber b than is needed in the second chamber b. This means that an excess of oil volume is promoted in the control pressure line 17. In this case, the flow compensating valve 23 is activated by the line 25b, whereby an excess of oil volume is prevented by oil volume flows from the control pressure line 16 via the pressure relief valve 11 of the feed system directly to the tank 5. Accordingly, it can be prevented with the quantity compensation valve 23 according to the invention that the pressures on the bottom side and on the rod side of the actuator 1 during operation of the Schnellsenkventils 7 equalize and affect the operation of the Schnellsenkventils 7 negative.

To set a response limit for a pressure difference between the rod and bottom side and for returning the flow compensating valve 23 to its original position centering springs may be provided, which are not shown in the drawing.

For the sake of clarity, a numerical example for k = 1, 21 is shown below, which corresponds to a typical ratio of the piston areas on the rod side and bottom side.

Case A: Lowering speed 1.2 times faster than quick-lowering valve

Q of cylinder (chamber a) 204 l / min

Q by pumps 2a, 2b 85 and 851 / min

Q by quick-lowering valve 34 l / min

Q_total (fast-flow pump 2b) 119 l / min Q requires 120 l / min from chamber b

-> 1 l / min is missing and is supplied by feed pump 6

Case B: Lowering speed 1.25 times faster than without quick-lowering valve

Q cylinder (chamber a) 212 l / min

Q by pumps 2a, 2b 85 and 851 / min

Q with quick-lowering valve 42 l / min

Q_total (quick-release valve + pump 2b) 127 l / min Q requires from chamber b 125 l / min

-> Surplus is forwarded through purge valve 23 via the pressure relief valve to the tank 5 shows a modified starting from the embodiment of FIG. 4 embodiment. In contrast to FIG. 4, a separate pressure limiting valve 11 'is provided here, which connects the quantity compensating valve 23 to the tank. The function is essentially identical to the function already described for FIG. 4. However, it is advantageous in the embodiment according to FIG. 5 that a function equal to a purge valve in a conventional hydrostatic drive is achieved via the separate pressure relief valve 11 '. This embodiment is therefore preferred. It also allows the pressure relief valve 11 'with respect to the quantities flowing through and the set pressure to adjust the quantity compensation function, without having to compromise on other functions of the pressure relief valve, such as a pressure limit of the allowable feed pressure to accept. A commonly existing pressure relief valve for limiting the feed pressure of a feed system is not shown explicitly in FIG. 5. However, it is usually present as well. In rare exceptional cases, a direct connection of the quantity compensation valve 23 with the tank volume could be provided.

The invention is not limited to the illustrated embodiment. Rather, advantageous combinations of the individual features are possible with each other.

Claims

claims

A hydraulic system, comprising: a hydraulic actuator (1) comprising at least a first chamber (a) and a second chamber (b), each having a connection for connection to the hydraulic system, at least one adjustable hydraulic pump (2) comprising at least one first port and a second port, wherein the first port is connected to the first chamber (a) of the hydraulic actuator and the second port is connected to the second chamber (b) of the hydraulic actuator (1), characterized in that the first chamber (A) is connected to the second chamber (b) via at least one fast-flow valve (7) with a controllable flow cross-section.

2. A hydraulic system according to claim 1, characterized in that the flow cross-section of the adjustable quick-lowering valve (7) by means of a control signal between 0 and 100% is adjustable.

3. Hydraulic system according to one of the preceding claims, characterized in that the controllable

Quick-lowering valve (7) when operating the hydraulic actuator (1) is closed against a load.

4. Hydraulic system according to one of the preceding claims, characterized in that a nominal size of the pump (2) is dimensioned after a required stroke speed.

5. Hydraulic system according to one of the preceding claims, characterized in that the hydraulic pump (2) is a dual-circuit pump.

6. Hydraulic system according to one of the preceding claims, characterized in that a memory

(4) for hydraulic fluid is connected to at least one pump (2,6).

7. Hydraulic system according to one of the preceding claims, characterized in that the quick-lowering valve (7) by means of an electro-proportional pilot valve (8) is actuated.

8. Hydraulic system according to one of the preceding claims, characterized in that a floating position valve (9) is provided, via which the first and the second chamber (a, b) of the actuator (1) with a tank volume (5) are connectable ,

9. Hydraulic system according to claim 8, characterized in that the float position valve (9) with a between the quick-flow valve (7) and the second chamber (b) formed line section is connected.

10. A hydraulic system according to claim 8 or 9, characterized in that the float position valve (9) is acted upon in the direction of a closed rest position with the pressure prevailing in the first chamber (a).

11. A hydraulic system according to claim 1 to 10, characterized in that the fast-lowering valve (7) is acted upon by the pressure of the first chamber (a) in the direction of a closed position.

12. A hydraulic system according to claim 1 to 11, characterized in that a flow compensating valve (23) is provided, via which the first or the second chamber (a, b) of the actuator (1) with a tank volume (5) preferably via a pressure relief valve (11, 11 ') are connectable.

13. A hydraulic system according to claim 12, characterized in that the quantity compensating valve (23) is designed such that an excess of hydraulic fluid during operation of the quick-lowering valve (7) preferably via a pressure relief valve (11, 11 ') to the tank volume (5) conduct.

14. Method for actuating a hydraulic actuator (1) with the following method steps:

- Generating a volume flow from a first chamber (a) of the hydraulic actuator (1) in a second chamber (b) of the actuator (1) and

Increasing the volume flow by opening a quick-lowering valve connecting the first chamber (a) to the second chamber (b). (7)

15. The method according to claim 14, characterized in that to increase the operating speed, first the pump (2) is adjusted to maximum delivery volume and then to further increase the operating speed, the controllable Schnellsenk- valve (7) in the direction of its open position.

16. The method according to claim 14 or 15, characterized in that the controllable quick-lowering valve (7) is closed,. when a load pressure prevailing in the first chamber (a) exceeds a threshold value.

17. The method according to claim 14 to 16, characterized in that an excess of the quick-flow valve (7) funded hydraulic fluid by means of a purge valve (23) to a tank volume (5) preferably via a pressure relief valve (11, 11 ') is promoted.