Abstract:
A device for compensating for hydraulic effective pressures in a hydraulic accumulator ( 9 ) and a hydraulic actuator ( 5 ) of a hydraulic system ( 11, 13 ) has a valve arrangement ( 27 ) for blocking a connection between the hydraulic actuator ( 5 ) and hydraulic accumulator ( 9 ) and has a control valve device ( 11 ) performing a pressure compensation when a predetermined difference in effective pressures is exceeded.

Description:
FIELD OF THE INVENTION 
     The invention relates to a device for compensating for hydraulic effective pressures in a hydraulic accumulator and a hydraulic actuator of a hydraulic system. 
     BACKGROUND OF THE INVENTION 
     In prior art hydraulic systems in which hydraulic actuators are used, for example, for support or lifting systems, hydraulic accumulators as spring or damper elements are hydraulically coupled to the actuator for cushioning or attenuating the movements of components moved by the hydraulic actuator. In some operating situations of such systems, however, an uncushioned, rigid dynamic connection between the actuator and the device actuated thereby is necessary, for example, for a hydraulically actuated boom intended to form a rigid support element, or for a tool to be controlled vibration-free when in use. In view of these requirements, the connection between the pertinent actuator and the hydraulic accumulator must be blocked. 
     In operation with the spring system blocked, the effective pressure in the hydraulic actuator changes according to the performance to be delivered by it. If at this point the system is transferred from the state of the blocked spring system back into the state with the hydraulic accumulator connected, a difference in the effective pressure between the hydraulic accumulator and the actuator leads to uncontrolled motion at the actuator. This uncontrolled motion poses a hazard to the system and a safety risk for system operators. 
     SUMMARY OF THE INVENTION 
     An object of the invention is to provide a device that prevents this safety risk from uncontrolled motion. 
     This object is basically achieved according to the invention by a pressure compensation device having a valve arrangement that blocks the connection between the hydraulic actuator and the hydraulic accumulator. The valve arrangement has an additional control valve that affects pressure compensation when a predetermined difference of the effective pressures is exceeded. This pressure compensation avoids the risk of uncontrolled motion when the system is transferred from the state of the blocked spring system into the state with the spring system released, because the respective effective pressures of the hydraulic accumulator and of the hydraulic actuator are matched to one another. 
     If, in the state of the blocked spring system, the pressure that is effective in the hydraulic accumulator is less than the effective pressure in the respective working situation in the hydraulic actuator, pressure compensation can easily take place in the conventional manner by the hydraulic actuator charging the hydraulic accumulator via a non-return valve up to a constant pressure. The non-return valve closes when the pressure is equal. 
     The particular advantage of the invention is that, when a higher pressure prevails in the hydraulic accumulator, this pressure is reduced by pressure drainage toward the tank side of the hydraulic system. 
     The valve arrangement can have a directional valve that, in its release state, establishes a direct fluid connection between the hydraulic actuator and the hydraulic accumulator and interrupts this fluid connection in its blocked state. The control valve can be activated depending on the transfer of the directional valve into the blocked state and can contain a drainage valve controllable by a difference of effective pressures that exceeds the preset value into the drainage valve release state in which a drainage path that reduces the pressure difference toward the tank side of the hydraulic system is formed. This arrangement ensures that the equalization of the effective pressures takes place not only by charging of the hydraulic accumulator, but that charging of the hydraulic accumulator can take place only up to a pressure level at which the prescribed pressure difference is not exceeded, because, when this pressure difference is reached, pressure compensation takes place via the drainage valve toward the tank side of the system. 
     The hydraulic actuator can have at least one lifting cylinder of a machine with a piston side producing the lifting force and with a rod side connected to a control block of the machine. The piston side of the lifting cylinder is connectable via the directional valve to the hydraulic accumulator. The control valve has a connection to the hydraulic accumulator and fluid paths to the piston side and to the rod side of the lifting cylinder. Two fluid paths contain non-return valves that clear the fluid path only to the side of the lifting cylinder carrying the higher effective pressure. 
     A drainage valve can be in the form of a pressure compensator. In the release state, the pressure compensator clears the drainage path toward the tank side from the connection to the hydraulic accumulator and from the fluid path cleared in each case and leading to the lifting cylinder. 
     To avoid generating noise or causing damage to the hydraulic accumulator, the drainage process can take place from the accumulator to the tank side only when the pressure difference is somewhat greater than zero. At the same time, preloading that intensifies the action of the closing pressure can be active on the pressure compensator. 
     The pressure compensator can have a slide valve piston that, for its displacement into the blocking position on one piston area, can be loaded both with the closing pressure from the hydraulic working circuit and loaded with the force of a preload spring. 
     Other objects, advantages and salient features of the present invention will become apparent from the following detailed description, which, taken in conjunction with the annexed drawings, discloses a preferred embodiment of the present invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Referring to the drawings which form a part of this disclosure: 
         FIG. 1  is a schematically simplified, side elevational view of a mobile machine in the form of a wheel loader, equipped with one exemplary embodiment of the device according to the invention; 
         FIG. 2  is a symbolic circuit diagram of the hydraulic system of the exemplary embodiment of the device according to the invention, shown in the operating state with the spring system released; 
         FIG. 3  is a circuit diagram of the device of  FIG. 2 , with the operating state being shown with an effective pressure in the hydraulic accumulator that is smaller than the effective pressure on the piston side of the lifting cylinder; 
         FIG. 4  is a circuit diagram of the device of  FIG. 2 , with the effective pressure in the accumulator being greater than on the piston side of the lifting cylinder; 
         FIG. 5  is a circuit diagram of the device of  FIG. 2  with the effective pressure on the rod side of the lifting cylinder being greater than on the piston side or in the hydraulic accumulator; 
         FIG. 6  is a functional and schematic side elevational view of a pressure compensator that serves as a drainage valve of the exemplary embodiment of  FIG. 2 ; 
         FIG. 7  is a symbolic representation of the pressure compensator of  FIG. 6 ; and 
         FIG. 8  is a side elevational view in section of a spool valve that serves as a pressure compensator of  FIG. 6  and that can be inserted into a valve block (not shown). 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
       FIG. 1  shows a mobile machine in the form of a wheel loader  1  with a shovel  3  coupled to a lifting cylinder  5 . The cylinder  5  forms the hydraulic actuator of the exemplary embodiment of the device according to the invention to be described. The piston side  7  of the lifting cylinder  5  produces the lifting force for the shovel  3  when pressure is supplied and is connected to a hydraulic accumulator  9 , indicated only symbolically in  FIG. 1 , via the hydraulic components not illustrated in  FIG. 1 . 
       FIGS. 2 to 5  in a symbolic representation show the circuit of the hydraulic system in different operating states.  FIG. 2  shows the state with the spring system released. A control block  13  of the machine (wheel loader  1 ) with a pressure supply (not shown), for controlled supply of the lifting cylinder  5  is connected to its piston side  7  and its rod side  15 . A valve arrangement  11  that forms the principal part of the hydraulic system has inputs or ports  17  and  19  connected to the piston side  7  and the rod side  15  of the lifting cylinder  5 , respectively. The hydraulic accumulator  9  and the tank  25  of the hydraulic system are connected to the outputs  21  and  23 , respectively, of the valve arrangement  11 . 
     As mentioned,  FIG. 2  shows the state of the released spring system. A directional valve  27  is in its release state as a result of its mechanical spring preload or spring  29 . The piston side  7  on the input or port  17  is connected directly to the hydraulic accumulator  9  at the output  21 , and the rod side  15  of the lifting cylinder  5  is connected via the input  19  directly to the tank  25  at the output  23 . In this operating state, the other hydraulic components are not involved in the operating process; i.e., the system effects a conventional cushioning/damping of the activity of the lifting cylinder  5 . 
     As mentioned, in certain operating situations a spring system is not useful or is detrimental. When a shovel  3  of a loader  1  is actuated, for example, spring compression or rebound has a negative effect on the accuracy of the positioning of the shovel  3 . The system is transferred into the state of the blocked spring system such that, by supplying a hydraulic control pressure via a control line  50 , the directional valve  27  is moved into the blocking state against the preload  29 , as detailed below. 
       FIGS. 3 to 5  illustrate three different operating modes for the spring system blocked in each case. In  FIG. 3  state, the piston side  7  of the lifting cylinder  5  is at a higher effective pressure than in the hydraulic accumulator  9 , as dictated by operation. Accordingly.  FIG. 3  shows with the thicker line the fluid connections that carry the higher pressure, specifically from the input  17  of the valve arrangement  11  to the blocked directional valve  27  via a line branch  31  and from the line branch  31  via a closing pressure control line  33  shown by the thick line to a control port  35  of a drainage valve  37 . This control port  35  is designated as the second control port. Corresponding to the effective pressure that prevails in the line branch  31  and that is higher than that in the line branch  39  indicated by the thin line at the input  19  and on the rod side  15  of the lifting cylinder  5 , a non-return valve  41  connected to the line branch  31  is opened so that the accumulator  9  at the output  21  is charged to the pressure of the piston side  7  via an accumulator line  43 . In this state, the non-return valve  45 , connected between the accumulator line  43  and input  19  in the same direction as non-return valve  41 , is closed. This arrangement of the non-return valves  41  and  45  causes the higher effective pressure from the inputs  17  and  19  to take effect in the system via a respective fluid path formed by opening of one or another non-return valve. Furthermore, in the connecting line to the accumulator  9  between the two port sites of the non-return valves  41  and  45  another non-return valve  46  is connected that, oriented toward the accumulator  9 , moves into its pertinent closed position. 
     Another control port  47  of the drainage valve  37 , referred to as the first control port, is connected via a control valve  49 , when it is in its opening state shown in  FIG. 3 , to the accumulator line  43  which in turn is connected to the input  17  or the input  19  corresponding to one or another fluid path, i.e., depending on which of the non-return valves  41  or  45  is opened. In the state shown in  FIG. 3 , the fluid path leads via the non-return valve  41  to the input  17  that carries the higher effective pressure. The pressure that prevails on the first control port  47  via the opened control valve  49  also serves as a hydraulic control pressure that hydraulically transfers the directional valve  27 , which directional control valve  27  is preloaded into the opening state by its spring preload  29 , into the closed state shown in  FIG. 3 , and thus, moves the entire system into the state of the blocked spring system. 
     With the released spring system in the state of  FIG. 2 , the control valve  49  is in its closed state caused by its actuating magnet  51  being energized so that the valve  49  is closed against its opening spring  52 . In this way, in the state of the released spring system, the first control port  47  of the drainage valve  37  and the control line  50  of the directional valve  27  are depressurized by connecting to the tank side  25 . The preload  29  therefore keeps the directional valve  27  in its opening state. If the power to the actuating magnet  51  is interrupted and the control valve  49  is opened, the directional valve  27  is hydraulically directed against its preload  29  into the blocked state via the control line  50 , and the system passes into the state of the blocked spring system, as is shown in  FIGS. 3 to 5 . 
     In the state shown in  FIG. 3 , in which the higher effective pressure prevailing in the line branch  31  charges the hydraulic accumulator  9  via the non-return valve  41  and the accumulator line  43 , on the first control port  47  and on the second control port  35  of the drainage valve  37  the same pressures prevail in each case, specifically via the control line  33  from the input  17  and via the opened non-return valve  41  and the opened control valve  49  likewise from the input  17 . The drainage valve  37  is a pressure compensator that is in the closed state when this constant pressure prevails on the control ports  47  and  35 . The drainage valve  47  in this closed state does not form a drainage path from the input port  53  to an output port  55  that leads via a drain line  57  by way of the output  23  to the tank  25 . Therefore, no drainage process takes place from the accumulator line  43  connected to the output  23  and the tank  25  via a pressure limitation valve  59  that forms an overpressure safeguard. A drainage valve  61  is likewise connected to the accumulator line  43  and that is manually opened only for maintenance purposes. 
       FIG. 4  conversely shows a state in which, likewise with the spring system blocked, the effective pressure in the hydraulic accumulator  9  is higher than the system pressure that is effective as dictated by operation on the piston side  7  of the lifting cylinder  5 , and thus, via the input  17  in the valve arrangement  11 . To illustrate this in  FIG. 4 , in the part uppermost in the figure, the accumulator line  43  is indicated by the thick solid line and in its lower line part by the thick broken line. The non-return valve  41  is closed corresponding to the effective pressure that prevails in the hydraulic accumulator  9 , which effective pressure is higher than in the lifting cylinder  5 . 
     The higher effective pressure of the hydraulic accumulator  9  is on the first control port  47  of the drainage valve  37  via the control valve  49  that is opened by the spring preload  52  and that is not energized. The second control port  35  carries the lower effective pressure of the input  17  via the line branch  31 . 
     As already mentioned, the drainage valve  37  has a pressure compensator shown symbolically in  FIG. 7  and in the form of an operating diagram in  FIG. 6 .  FIG. 8  shows a longitudinal section of one practical embodiment. Drainage valve  37  is a spool valve with slide valve piston  65  axially displaceable in the valve housing  63 , shown in the closed position. This closing is caused by a hydraulic closing pressure that acts on the second control port  35 , amplified by a mechanical preload force  67  in  FIGS. 6 and 7 . The drainage valve  37  opens by a hydraulic opening pressure that is active on the first control port  47 , assuming that the opening pressure on the slide valve piston  65  causes a higher opening pressure than the closing pressure that prevails on the control port  35 , amplified by the preload force  67 . In other words, the condition for the drainage valve  37  to open to form a drainage path from the input port  53  to the output port  55  and thus to the tank  25  is when the closing forces acting on the slide valve piston  65  resulting from the pressure on the second control port  35 , plus the mechanical preload  67 , is smaller than the opening pressure produced by the hydraulic pressure on the first control port  47 . Therefore
 
 F   preload   +F   pressure35   &lt;F   pressure47  
 
     In the state depicted in  FIG. 4 , the pressure from the hydraulic accumulator  9  is drained until only a given, desired low pressure excess between the accumulator  9  and thus the input port  53  remains relative to the control port  35 , i.e., the lifting cylinder  5 , corresponding to the design of the pressure compensator that forms the drainage valve  37 , specifically the effective piston areas and the effective preload force  67 . This state means that a drain process cannot lead to reducing the pressure in the hydraulic accumulator  9  to a value of zero. 
     Advantageously, the opening pressure difference dictated by the piston geometry and the preload force  67  can be a pressure level of approximately 8 bar.  FIG. 8  shows two helical springs  69  and  71  acting on a two-part slide valve piston  65  for producing the preload force  67  and preloading the piston  65  into the illustrated closing position to the right in the figure, in which the input port  53  located on the axial end of the spool housing  53  on the right side in the figure is blocked relative to the output port  55 . In addition to the preload force  67 , the hydraulic pressure from the second control port  35  acts on the side of the piston  65 , which side is the left one in the figure. As the opening pressure for moving the piston  65  in the figure to the left, the right piston area is subjected to the opening pressure via the first control port  47 . 
     To ensure that the pressure present on the input port  53  does not take effect as the effective control pressure that determines the behavior of the pressure compensator, the piston area  73  indicated in  FIG. 6  and bordered by the control edges  75  and  77  between the ports  53  and  55  importantly be considerably smaller than the effective piston areas  79 ,  79   a , and  81  on the pressure spaces on the control port  47  or control port  35 . 
       FIG. 5  relates to another state in which, at the input  19  of the valve arrangement  11 , the higher effective pressure prevails, compared to the pressure at the input  17  or the pressure in the hydraulic accumulator  9 . This operating state arises when a device runs up against an obstacle during operation of a machine with the spring system blocked. This state can be the case, for example, when a mobile device, such as a wheel loader  1 , with its shovel  3  runs up against an obstacle that forms an elevation. As a result of this situation, the weight of the wheel loader  1  resting on the shovel  3  pushes the piston of the pertinent lifting cylinder  5  into the rod side  15 , causing an overpressure to form on the rod side  15 . This overpressure takes effect via the input  19 , with the non-return valve  45  opening in this state, as well as via the opened control valve  49  on the first control port  47  of the drainage valve  37 . When the opening condition is met, i.e., a higher pressure on the port  53  compared to the control port  35  connected to the input  17  via the line branch  31 , the drainage valve  37  then opens. As a result of valve  37  opening, in turn the drainage path to the tank  25  is opened, causing the pressure of the accumulator line  43  to be relieved. The higher pressure in the control port  47  ensures that the valve  37  is not in the blocking position. 
     As  FIGS. 6 and 8  show in particular, the actual pressure compensator is formed by the helical spring  69  and by the effective pressure surfaces of the axially displaceable slide valve piston  65 . The blocking piston made as a valve spool is in turn formed by the helical spring  71  and the effective piston area  81  of the indicated blocking piston part. 
     The piston  65  in  FIG. 6  can be made in several parts to form a non-return valve, i.e., the multipart design prevents opening of the valve seat  55  and unwanted backflow of the fluid into the system when a pressure prevails on the port  55  that is higher than that pressure formed by the preload forces of the helical springs  69  and  71  plus the effective compressive force by the pressure on the second control port  35 . If this non-return valve function is to be omitted, the illustrated slide valve piston arrangement can also be made in one piece (not shown). 
     The invention thus ensures that the safety function is pressure compensation for all operating modes. The construction of the drainage valve  37  as shown in  FIGS. 6 and 8  is not mandatory. Any valve construction whose operation corresponds to the aforementioned opening and closing conditions can be used. The construction of the two-part slide valve piston  65  depicted in  FIG. 8  and the construction of the piston part to the right in this figure at the input port  53  forming a non-return valve loaded by the spring  69  with low closing force are not mandatory. In this construction, the closing spring  71  forms the principal part of the preload  67  in  FIGS. 6 and 7  and amplifies the closing force of the valve.