Abstract:
The invention concerns a hydraulic control ( 1 ) with a supply connection arrangement ( 7 ) having a high-pressure connection (P) and a low-pressure connection (T), a working connection arrangement having two working connections (A, B) connectable with a consumer, a control valve ( 8 ) with a valve element ( 9 ) between the supply connection arrangement and the working connection arrangement and a compensation valve ( 11 ), which is located between the high-pressure connection (P) and the control valve ( 8 ) and is acted upon in the closing direction by a pressure between the compensation valve ( 11 ) and the control valve ( 8 ). 
     It is endeavoured to ensure the most favourable energy consumption possible. 
     For this purpose, in the opening direction the compensation valve ( 11 ) is acted upon by a pressure of a selection device ( 29, 30, 30′, 38 ), which optionally supplies the compensation valve ( 11 ) with a pressure control pressure or a flow control pressure.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   Applicant hereby claims foreign priority benefits under U.S.C. § 119 from German Patent Application No. 10 2004 063 044.5 filed on Dec. 22, 2004, the contents of which are incorporated by reference herein. 
   FIELD OF THE INVENTION 
   The invention concerns a hydraulic control with a supply connection arrangement having a high-pressure connection and a low-pressure connection, a working connection arrangement having two working connections connectable with a consumer, a control valve with a valve element between the supply connection arrangement and the working connection arrangement and a compensation valve, which is located between the high-pressure connection and the control valve and is acted upon in the closing direction by a pressure between the compensation valve and the control valve. Further, the invention concerns a method of controlling a hydraulic consumer, which is controlled by a control valve in a pressure control operation mode. 
   BACKGROUND OF THE INVENTION 
   Such a hydraulic control and such a method are known from DE 198 00 721 A1. In the opening direction, the compensation valve is acted upon by a spring and a pressure, which can be supplied via a fixed throttle. The fixed throttle is part of a pressure divider between the outlet of the compensation valve and the low-pressure connection, which here is a tank connection. Thus, the compensation valve ensures a pressure control, in which the motor inlet pressure has a value, which is substantially determined by the position of the control valve. 
   In the return pipe from the motor to the low-pressure connection, a compensation valve and a load-retaining valve are arranged in series. Via a pilot pipe the load-retaining valve is supplied with the motor inlet pressure in the opening direction and via a further pilot pipe with the pressure at the outlet of the load-retaining valve. Thus, under the influence of a spring, the load-retaining valve adjusts so that it does not open until the pressure difference has overcome the spring force. 
   When, now, this motor is lowered under a load, a relatively high inlet pressure is required. For example, the control valve slide has to be opened relatively much, and, in dependence of the design, a larger or smaller slide movement is required to control the high pressure. This is energetically unfavourable, as this high pressure merely has to be available for opening the load-retaining valve. 
   Another possibility of using the compensation valve is shown in DE 102 16 958 B3. Here, the compensation valve is controlled by a pressure difference over the control valve and keeps the pressure difference over the control valve constant. In this manner, a flow control is realised, in which the amount supplied to the consumer depends on the position of the valve element. The more the valve element is displaced, the larger are the inlet flow and the outlet flow. 
   U.S. Pat. No. 4,981,159 shows a hydraulic control, which can be used with different valve elements as pressure control on the one side and as flow control on the other side. For this purpose, the valve element, which also has the form of a slide, merely has to be replaced. In principle, such a replacement is not difficult. However, it can only be made, when the system is pressureless or, even better, empty. Thus, a change of operation modes still requires certain efforts. 
   BRIEF SUMMARY OF THE INVENTION 
   The invention is based on the task of providing the most favourably energy consumption pattern. 
   With a hydraulic control as mentioned in the introduction, this task is solved in that in the opening direction the compensation valve is acted upon by a pressure of a selection device, which optionally supplies the compensation valve with a pressure control pressure or a flow control pressure. 
   With this embodiment, it is possible to operate the hydraulic control optionally in a pressure control operation mode or a flow control operation mode. It is not necessary to make any alteration. It is sufficient to use different pressures, which are selected via the selection device and then specifically supplied to the compensation valve. Thus, it is possible to select the pressure control pressure or flow control pressure, which permits the most favourable energetic operation mode. The selection device can be provided for both movement directions of the consumer. In many case, however, it will be sufficient to provide the selection device for only one movement direction, in which negative loads can occur. Further, with this embodiment, a substantially more comfortable operation of the control, can be achieved. When, until now, it has been desired to lower a negative load, for example to collapse a crane jib, first a negative load and then a positive load had to be supplied to ensure a complete collapse of the crane jib. For this purpose, an actuating element of the control had to be moved to manage the transition from the negative to the positive load. With the new embodiment, the actuating element, for example a handle, can be left in a set position, and the control will automatically change to flow control, when the force gets positive. 
   It is preferred that the selection device supplies the higher of the pressures, pressure control pressure and flow control pressure, to the compensation valve. This has two advantages. Firstly, it is easier to decide, which of the two pressures should be chosen. Secondly, also the operation of the selection device can be automated in this manner. 
   Preferably, an actuation of the control valve from a predetermined position will make the selection device pass on firstly the pressure control pressure and secondly the flow control pressure to the compensation valve. The position mentioned can, for example, be a “zero position” or “neutral position”, which is used as an example in the following explanation. Depending on the design of the control valve, this predetermined position can, however, also be somewhere else. When the control valve is moved from its zero position, it opens increasingly and thus passes on hydraulic fluid from the high-pressure connection, which is usually made as pump connection, to a working connection. In the initial phase of this opening section, the control is then operated in a pressure control operation mode, in which the pressure at the outlet of the control valve substantially depends on the position of the valve element of the control valve. Of course, the individual pressures depend on the exact design of the valve element, for example a valve slide. Thus, here the explanation has to be understood as an example. It merely serves a better understanding of the invention. This pressure can then, for example, be used to open other valves of the control, for example a load-retaining valve. This load retaining valve then merely has to be dimensioned for this relatively small pressure, which is enabled by the pressure control. It is also possible to act oppositely and first select a load-retaining valve and then dimension the remaining system. When this minimum pressure is exceeded, the selection device automatically switches to a flow control operation mode. In a flow control operation mode the pressure is then determined practically exclusively by the consumer, that is, only the absolutely necessary pressure is provided. The control valve, which is preferably a proportional valve, then supplies the corresponding amount of hydraulic fluid, that is, to put it simply, it controls the speed, with which the consumer is driven. Thus, with this embodiment the energetically most favourable pressure, that is, the pressure required by the consumer, is set in a pressure area, which is limited downwards by the minimum pressure specified by the pressure control and upwards, if required, by an overpressure valve. Thus, in the end, the external conditions determine the form of control to be active. Of course, this also applies in the “initial phase”. 
   Preferably, the selection device is on the one side connected with a working pipe located between the control valve and a working connection and on the other side with a control pipe connected with a load-sensing pipe. Of course, this applies, when the control valve is in the operation state, that is, the valve element has been deflected from its resting position and has created a connection between the compensation valve and one of the working connections. The actuation of the valve element increases the pressure in the working pipe. As long as this pressure is smaller than the pressure in the control pipe, a pressure control occurs. During the pressure control, the pressure at the working connection is substantially depending on the position of the valve element. If the valve element is further activated, the pressure at the working connection will, depending on the external conditions, at some time exceed the pressure in the control pipe. In this case, a flow control occurs, in which the pressure at the working connection is determined by the pressure of the consumer. Thus, an energetically extremely favourable operation can be realised, as only the pressure required to drive the consumer has to be supplied. In the control pipe there is, in a manner of speaking, an “artificial load signal”. 
   Preferably, the control pipe is connected with an outlet of a pressure divider, which is located between the compensation valve and the low-pressure connection. The same pressure divider can also be used to generate the load-sensing signal. However, usually a further throttle is located between the pressure divider and a load-sensing connection (LS connection), which throttle causes a certain decoupling. The outlet of the pressure divider supplies a pressure, which acts upon the compensation valve in the opening direction. This is a relatively simple manner of providing the pressure control. 
   Preferably, the pressure divider has at least two throttles, of which one can be adjusted by the valve element of the control valve. This throttle is usually the throttle located between the outlet and the low-pressure connection. 
   In a preferred embodiment, the pressure divider has two throttles, which can both be adjusted by the valve element of the control valve. When the throttles of the pressure divider have a constant value, the pressure at the outlet of the control valve remains substantially constant in the pressure control area. When these throttles have a variable value, the pressure can be increased or reduced. 
   In a preferred embodiment, the selection device has a non-return valve, which opens in the direction of the compensation valve. This is a relatively simple embodiment, which is, however, sufficient, when merely the higher of the two pressures has to be passed on to the compensation valve. 
   It is preferred that the non-return valve is located in the valve element of the control valve. In this case only few modifications of the control itself are required. Merely a small modification in the valve element of the control valve is required. 
   The selection device can also comprise a shuttle valve. In a manner of speaking, a shuttle valve is a non-return valve with two non-return valve functions. Also such a shuttle valve can be located in the valve element of the control valve. 
   Preferably, a load-retaining valve is located at least one working connection, which load retaining valve can be opened by the pressure at the other working connection. Such a load-retaining valve is also called “overcenter” valve. A predetermined opening pressure is required for such a load-retaining valve. This opening pressure cannot be made too small, to prevent the load-retaining valve from opening unintentionally, when leakages or other unfavourable conditions lead to a pressure build-up, which causes the opening of the load-retaining valve. With a pilot control device, the opening pressure of the load-retaining valve can now be kept relatively high, thus keeping the required safety distance to pressures building up parasitically without having to drive the energetic efforts for opening the load-retaining valve too high. To open the load-retaining valve, a pressure merely has to be built up at the other working connection, which is sufficient to activate the pilot control device. Such a pressure can, for example, correspond to the minimum pressure specified by the pressure control. Thus, to lower a load only the absolutely necessary pressure has to be built up. This pressure can, for example, correspond to the pressure of the opening spring at the compensation valve plus the pressure at the outlet of the pressure divider before the control valve. Of course, in another such embodiment it is also possible to use a return compensation valve between the consumer or the working connection and the control valve. 
   It is preferred that the pilot control device has a pilot valve element controllable by the pressure at the other working connection, said pilot control device making in the controlled state a connection from one working connection to a control inlet of the load-retaining valve and interrupting it in the uncontrolled state. This is a relatively simple design of a pilot control device. 
   Preferably, the working connection arrangement is connected with an anti-cavitation device, which has an anti-cavitation valve with an anti-cavitation valve element, which is displaceable by means of a pressure at a working connection and creates a connection between a consumer connection and the other working connection. The connection can be realised in that in the direction of the consumer practically no restrictions exist in the form of throttles, narrow passages in a valve block or the like. Accordingly, the refilling can take place at a lower pressure than before, so that also a pushing operation, that is, an operation with negative loads, will also require relatively less additional energy. 
   Preferably, the outlet of the selection device is connected with a pressure limitation valve. Via the pressure limitation valve, which is set in dependence of the application, for example, the pressure control pressure can be increased or decreased with the change of position of the valve element of the control valve. 
   The task is solved with a method as mentioned in the introduction in that the control valve alternatively controls the consumer in a flow control operation mode and that the switching between the pressure control operation mode and the flow control mode occurs automatically in dependence of the ruling pressures. 
   Thus, it is possible to operate the consumer in an energetically favourable area. In the flow control operation mode the pressure of the consumer is determining. In the pressure control operation mode the pressure of the control valve is determining. The switching between these two operation modes then depends on the pressures at the consumer connection. For example, the selection device mentioned above can be used for this purpose. However, such a method can also be realised otherwise, for example with electrically controlled components. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     In the following, the invention is described by means of preferred embodiments in connection with the drawings, showing: 
       FIG. 1  a first embodiment of a hydraulic control; 
       FIG. 2  a schematic view explaining the pressure conditions; 
       FIG. 3  a second embodiment of the hydraulic control; 
       FIG. 4  a simplified view of a further embodiment of the hydraulic control; 
       FIG. 5  an embodiment modified in relation to  FIG. 4 ; 
       FIG. 6  an embodiment modified in relation to  FIG. 4 ; 
       FIG. 7  a schematic view of a consumer with a load-retaining valve; and 
       FIG. 8  a schematic view of an anti-cavitation device. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
     FIG. 1  shows a hydraulic control  1  for the control of a consumer  2 , here a piston cylinder arrangement with a piston  3  and a cylinder  4 . The piston  3  divides the cylinder into a first pressure chamber  5  and a second pressure chamber  6 . The two pressure chambers  5 ,  6  are connected with working connections A, B of the control  1 . Together, the two working connections A, B form a working connection arrangement. 
   The control  1  has a supply connection arrangement  7 , which has a high-pressure connection P in the form of a pump connection, a low-pressure connection T in the form of a tank connection and a load-sensing connection LS. 
   Between the supply connection arrangement  7  and the working connection arrangement A, B is located a control valve  8 , which has a valve slide  9  as valve element. By means of a merely schematically shown actuator  10 , for example in the form of an electromagnetic actuator or a pilot controlled actuator, the valve slide  9  can be displaced to a total of five different operation modes. These operation modes are shown by means of five positions a to e. Actually, however, the valve slide  9  of the control valve  8  is practically continuously movable, so that it can assume practically any intermediate position. Here, the control valve  8  is a proportional valve. 
   In a manner known per se and therefore not described in detail, the valve slide  9  has grooves and other recesses, if required bores and the like, on its circumference, which overlap corresponding annular grooves, recesses and bores in a housing of the control valve  8 , thus releasing or blocking in a more or less throttled manner certain connections between the supply connection arrangement  7  and the working connection arrangement A, B in dependence of the position of the valve slide  9 . Examples showing the housing of such control valves and a corresponding slide are, for example, known from U.S. Pat. No. 4,981,159 mentioned in the introduction. Depending on the requirements, a person skilled in the art will be able to make such a slide and a corresponding housing. 
   A compensation valve  11  is located between the control valve  8  and the high-pressure connection P. In the opening direction the compensation valve is loaded by the force of a spring  12  and the pressure in a control pipe  14 . In the closing direction the compensation valve  11  is connected via a pipe  13  with its outlet, that is, a point between the compensation valve  11  and the control valve  8 . Thus, in the closing direction the inlet pressure of the control valve  8  acts upon the compensation valve  11 . 
   For reasons of simplicity the working connection A is in the following called “lifting connection”, as through this connection hydraulic fluid is supplied to the larger pressure chamber  5 , which leads to a lifting or extension of the piston  3 . The working connection B, however, is called “lowering connection”. Here pressurised hydraulic fluid must be supplied to lower or retract the piston  3  again. A load-retaining valve  15  is connected with the lifting connection A, which load-retaining valve  15  can be opened by the pressure at the lowering connection B. The load-retaining valve  15  is bridged by a non-return valve  16  opening in the direction of the first pressure chamber  5 . 
   The lifting connection A is connected via a return compensation valve  17  with a first working outlet  18  of the control valve  8 . The control valve  8  has a second working outlet  19 , which is connected with the lowering connection B. When negative loads occur, the lifting connection A is controlled by the return compensation valve  17 , as known from, for example DE 102 16 958 B3. 
   Further, the control valve  8  has a first load-sensing outlet  20  and a second load-sensing outlet  21 . In the shown neutral position c of the valve element  9 , the first working outlet  18 , the second working outlet  19 , the first load-sensing outlet  20  and the second load-sensing outlet  21  are connected with the low-pressure connection T. Thus, in a manner of speaking, the consumer  2  is in a “floating position”. 
   Located next to the neutral position c are blocking positions b, d of the valve element  9 , in which merely the two load sensing outlets  20 ,  21  are connected with the low-pressure connection T. The two working outlets  18 ,  19 , however, are blocked. In all three positions b, c, d mentioned until now, a pressure inlet  22  of the control valve  8  is blocked. The pressure inlet  22  is connected with the outlet of the compensation valve  11 . 
   In a lifting position e the valve slide  9  is displaced so that the first working connection  18  and the first load-sensing outlet  20  are connected with the pressure inlet  22 . The second pressure outlet  19  and the second load-sensing outlet  21  are connected with the low-pressure connection T. Pressurised hydraulic fluid is then supplied to the lifting connection A and reaches the pressure chamber  5  via the non-return valve  16 . The piston  3  moves to the right. This is so to speak a normal operation mode. 
   In a lowering position a, however, the second working outlet  19  is connected with the pressure inlet  22 , while the first working outlet  18  and the first load sensing outlet  20  are connected with the low-pressure connection T. 
   The second load-sensing outlet  21  is connected with an outlet  23  of a pressure divider, which is formed by two throttles  24 ,  25 . The throttle  25  is located between the outlet  23  and the low-pressure connection T. The throttle  24  is located between the outlet  23  and the pressure inlet  22 . The throttle  24  can be a constant throttle, whose flow resistance is independent of the position of the valve slide, whereas the flow resistance of the throttle  25  is variable by means of adjustments of the valve slide  9 . Via a bleed  26  and a shuttle valve  27  the second load sensing outlet  21  is connected with the control pipe  14 . Further, the second load sensing outlet  21  is connected with the load sensing connection LS of the supply connection arrangement  7  via a second shuttle valve  28  connected in series with the shuttle valve  27 . 
   The first shuttle valve  27  is connected with the first load sensing outlet  20  via a bleed  26   a.    
   The second load sensing outlet  21  is connected with an inlet of a selection device  29 . Also the second working outlet  19  is connected with this selection device. The selection device  29  has a non-return valve  30  in the pipe connected with the second working outlet  19 , so that the larger of the two pressures at the second working outlet  19  and the second load sensing outlet  21  is always available at the outlet  31 . 
   This has the following effect: When the valve slide  9  is displaced to its lowering position a, the lowering outlet B is supplied with pressure. At the same time, the pressure at the lowering outlet B opens the load-retaining valve  15 , so that pressurised hydraulic fluid can escape from the pressure chamber  5 . The compensation valve  11  is controlled in two different manners, again depending on the external conditions. This is explained by means of the following example: 
   Initially, the pressure at the second load-sensing outlet  21  is larger than the pressure at the second working outlet  19 . The reason is that at the beginning of its movement the valve slide  9  causes a relatively large throttling effect with the control valve  8 . In this case, the pressure at the second working outlet  19  changes proportionally with the movement of the valve slide  9 . This is shown as a section P 1  in  FIG. 2 . In this area the control  1  works as a pressure control. However, as soon as a further movement of the valve slide  9  causes a reduction of the throttling effect between the valve slide  9  and the housing of the control valve  8 , and the pressure at the second working outlet  19  increases over the pressure at the second load-sensing outlet  21 , this pressure is used for controlling the compensation valve  11  and the control valve  8  works as a flow control valve, that is, the flow is now set in dependence of the position of the valve slide  9  in the control valve  8 . The pressure, however, is determined by the consumer  2 . The upper limit is fixed by an overpressure valve  32 . A corresponding overpressure valve  32 ′ is also mounted at the other working connection A. 
   When the throttle  24  between the pressure inlet  22  and the outlet  23  is also made to be variable, that is, changes with the position of the valve slide  9 , this result in the lower ramp  33  shown in  FIG. 2 , which shows the minimum pressure of the control valve in dependence of the deflection x of the slide. At the top in  FIG. 2  is shown a hybrid pressure H, that is, a pressure which is combined partly by the pressure control and partly by the flow control. The area “FC control” shows that here only the flow is controlled. The pressure adjusts automatically. When the external conditions are different, also other sequences of the pressure and flow control can occur. 
   In a manner known per se, a pilot-controlled stop valve  34  is also allocated to the lowering connection B. 
   By means of  FIG. 4 , the mode of functioning shall be explained once again. Same parts are provided with the same reference numbers. Further shown is a variable pump  35 , which is controlled via the load-sensing connection LS. The control valve  8  is here merely symbolised by two “large” throttles  36 ,  37  and the “small” throttle  25  as well as the throttle  24 . The large throttles  36 ,  37  and the small throttle  25  are adjustable in dependence of the position of the valve slide  9  in the control valve  8 . 
   When the valve slide  9  is displaced in the control valve  8 , the throttles  36 ,  37  open and the throttle  25  closes. This leads to the increasing curve for the minimum pressure shown in  FIG. 2 . When the throttle  25  opens, a falling curve occurs. When the throttle  36  is still slightly open, that is, provides a large resistance, then, in dependence of the external conditions, that is, the other pressures in the system, for example the pressure at the second working outlet  19  is smaller than the pressure at the pressure inlet  22 . Over the fixed throttle  24  only a small pressure drop occurs, as at the beginning of the movement of the valve slide  9  the variable throttle  25  is only slightly opened. Accordingly, the pressure at the outlet  23  is higher than the pressure at the second working outlet  19 , and the non-return valve  30 , which can, as shown, also be located in the valve slide  9 , remains closed. Thus, the compensation valve  11  is controlled by the pressure difference between the pressure inlet  22  and the outlet  23 . The pressure at the second working outlet  19  is then proportional to the displacement of the valve slide  9 . The pressure is dimensioned so that, at least when it has reached its maximum value, it is sufficient to open the load retaining valve  15 . A higher pressure is not required to open the load retaining valve  15 . In this area the valve slide is moved by approximately 1 to 2 mm. 
   When, now, the throttling resistance of the throttle  36  further decreases, the pressure at the second working outlet  19  increases until it exceeds the pressure at the outlet  23 . In this case, the non-return valve  30  opens, that is, the selection device  29  switches from the pressure control to the flow control. As soon as the non-return valve  30  has opened, the flow to the consumer  2  is determined by the position of the valve slide  9 . The pressure, however, is determined by the consumer. In this area the valve slide is moved by a further 3 to 4 mm. 
   This gives an extremely energy-saving operation. A corresponding operation diagram is shown in  FIG. 4   a . At least a minimum pressure H 1  is reached. This minimum pressure is defined by the pressure division between the throttles  24  and  25 . A maximum pressure H 2  is limited by the overpressure valve  32 . Between H 1  and H 2  the pressure through the consumer  2  is determined. 
     FIG. 5  shows a modified embodiment. Same elements have the same reference numbers. The non-return valve  30  is replaced by a shuttle valve  38 , whose one inlet is connected with the second working outlet  19  and whose other inlet is connected with the outlet  23 . As can be seen from  FIG. 5   a , practically the same operation behaviour occurs here. The shuttle valve  38  passes on the higher of the two pressures from the second working outlet  19  and the outlet  23  to the compensation valve  11 . 
   If required, also the shuttle valve  38  can be integrated in the valve slide  9 . 
     FIG. 6  is a schematic view of an embodiment, which substantially corresponds to the embodiment in  FIG. 4 . Here, the control pipe  14  is not only connected with the outlet  23 , but additionally with a relief valve  39 , which opens in the direction of the tank T. The relief is set in dependence of the consumer  2 . As shown in  FIG. 6   a , this causes a minimum pressure curve  40  in the flow control area, which can be displaced between two limits  41 ,  42 . 
   In all three embodiments the pressure during flow control is determined by the consumer  2 . When the pressure supplied by the pressure control is too small to move the consumer, for example a load, the flow control takes over. 
   During the pressure control a minimum pressure occurs, which is determined by the throttle  24 . This minimum pressure is set so that it is sufficient to open the load-retaining valve  15 . One possibility of reducing this pressure at the lowering connection B will be discussed below in connection with  FIG. 7 . 
   In  FIG. 1  the control is designed so that it can activate a motor for lifting a load. Accordingly, it is sufficient for the selection device  29  to have a non-return valve  30  only for the lowering connection B. 
     FIG. 3  shows a control  1 , which is meant for driving a consumer  2 , which can be activated in both directions and which can also provide a negative load in both directions, for example during a pushing operation in connection with forward or backward driving of a rotary motor driving a vehicle. 
   The same parts have the same reference numbers as in  FIG. 1 . 
   The most essential difference in relation to  FIG. 1  is that a non-return valve  30 ,  30 ′ is now provided for each of the two working outlets  18 ,  19 , so that the compensation valve  11  can cause both a pressure control of the control valve  8  and a flow control in each movement direction. Accordingly, also a pressure divider with two throttles  24 ′,  25 ′ and an outlet  23 ′ are provided for the second working outlet A, the outlet  23 ′ being connected with the bleed  26   a , when the valve slide  9  is moved to the position E. The two blocking positions b, d are not provided here. 
   When the valve slide  9  is in the position e, the non-return valve  30 ′ in a manner of speaking decides, if the pressure at the first working outlet  18  or at the first load-sensing outlet  20  is higher, and should be used for controlling the compensation valve  11  via the control pipe  14 . 
   When, now, only the lowest possible pressure always rules at the lowering connection B, it could of course be difficult to open the load-retaining valve  15 . Means for this are shown in  FIG. 7 . 
   The load-retaining valve  15  has a control inlet  43 , which is connected with a pilot control device  44 . The pilot control device has a slide  45 , which can be displaced under the effect of a pressure at the lowering connection B. In the shown, non-displaced position the control inlet  43  of the load-retaining valve  15  is practically short-circuited or connected with the low-pressure connection T. 
   When, now, the pressure at the lowering connection B increases to a predetermined value, the slide  45  is displaced and connects the pressure chamber  5  with the control inlet  43  via a shuttle valve  46 . In this case, the load-retaining valve  15  is opened. At the same time, only small pressures are required at the lowering connection B. 
   In a transmission drive  2 ′ the pushing operation requires a refilling of hydraulic fluid to prevent cavitation. To enable this refilling at low pressures,  FIG. 8  shows an anti-cavitation device  47 , which can be connected with the two working connections A, B. Of course, further elements can be located between the anti-cavitation device  47  and the control  1 , for example the load-retaining valve  15  shown. 
   By means of throttles  48 ,  49  resistances are shown which can occur because of valve characteristics in a valve block, which is not shown in detail, with which the drive  2 ′ is connected. 
   The drive  2 ′ is connected with both working connections A, B. Further, it is connected with a common supply point  52  via two non-return valves  50 ,  51 . In this connection, the non-return valves  50 ,  51  open in the direction of the drive  2 ′. 
   The supply point  52  is connected with the outlet  53  of an anti-cavitation valve  54 . The anti-cavitation valve  54  has a slide  55 , which is acted upon by a control pressure from both working connections A, B. If the pressure at the working connection A is larger than the pressure at the working connection B, the slide  55  is displaced so that the working connection B is connected with the outlet  53 . The drive  2 ′ can then suck hydraulic fluid with lower pressure from the working connection B. This working connection will usually be connected with the tank. 
   In the opposite case, the pressure at the working connection B pushes the slide  55  so that the outlet  53  is connected with the working connection A, and the drive  2 ′ can then suck hydraulic fluid with lower pressure from the working connection A. 
   As the supply takes place after the throttles  48 ,  49  and thus occurs with relatively small resistances, only a relatively low pressure is required for the refilling. When until now approximately 50 bar have been required for the refilling to consider the throttling losses at the throttles  48 ,  49  (which are parasite losses), now, for example, 30 bar will be sufficient. 
   With the control, a load is possible, which is smaller than a set value of, for example, 30 bar. Over this load there is then a control according to the load level, which is specified by the consumer, in other words, a flow control. 
   The control permits a meter-in function or a meter-out function, respectively, the system itself selecting the possibility to be used. 
   With negative loads, a transmission drive  2 ′ can always provide a positive pressure at the inlet to protect against cavitation. In a cylinder application ( FIG. 1 ) is can be ensured that by means of the defined minimum pressure the load-retaining valve is rendered non-functional, that is, can be opened, when the load is negative. Also here there will be practically no cavitation. 
   While the present invention has been illustrated and described with respect to a particular embodiment thereof, it should be appreciated by those of ordinary skill in the art that various modifications to this invention may be made without departing from the spirit and scope of the present invention.