Patent Publication Number: US-2006011875-A1

Title: Hydraulically controlled valve comprising at least one hydraulic drive unit

Description:
The invention pertains to a hydraulically controlled valve with at least one hydraulic drive according to the introductory clause of claim  1 .  
      A load-holding brake valve which can be controlled by a hydraulic drive is known from WO 97/32136 A1. The main piston of the load-holding brake valve is actuated by the plunger of a control piston. A control pressure moves this control piston against the pressure of a control spring. These types of load-holding brake valves are suitable for driving double-acting hydraulic consumers, for example, which are subject to mechanical loads. Depending on the type of mechanical load, such devices tend to oscillate. Arrangements such as cranes with very long lift arms, for example, are known. As the result of an impact, for example, an oscillation can be caused, which causes the volume flow rate of the hydraulic oil to fluctuate. Oscillations can also originate in the hydraulic system itself, however, when the control of a movement is begun and/or a movement is accelerated or delayed. As a result of such oscillations, the speed of the hydraulic consumer is no longer constant, which means in turn that it becomes difficult or impossible to control such movements precisely.  
      A directional control valve which is suitable for driving double-acting hydraulic consumers is known from WO 02/075162 A1. It is disclosed here that the slider piston of the directional control valve can be moved by at least one drive. A solution with two hydraulic drives is shown. A drive piston which can be moved by a control pressure against a spring is provided in each of these drives. This drive piston can, for example, move the slider piston of the directional control valve by way of a piston rod. It is also possible for oscillation problems to occur in these types of arrangements.  
      A hydraulic, directly-controlled pressure-limiting valve of the sliding type is known from DE 24 31 785 A1. Because the differential piston which is present is controlled directly, this valve does not have a hydraulic drive.  
      A valve which can be used as a pressure-limiting valve is known from US  2 , 361 , 881  A. This valve does not have a hydraulic drive either.  
      A spring-loaded pressure relief valve, which also lacks a hydraulic drive, is known from DE-AS 1 254 925.  
      The invention is based on the task of creating a valve which is hydraulically controlled by at least one hydraulic drive and which is insensitive to both externally and internally induced oscillations without any impairment to the response sensitivity.  
      The task indicated above is accomplished according to the invention by the features of claim  1 . Advantageous elaborations can be derived from the dependent claims. 
    
    
      Exemplary embodiments of the invention are explained in greater detail below on the basis of the drawing:  
       FIG. 1  shows a diagram of the details essential to the invention on the basis of an example of a load-holding brake valve;  
       FIG. 2  shows a diagram, not to scale, of a part of a control piston in a primary control pressure chamber;  
       FIGS. 3   a - 3   c  show hydraulic diagrams of the various operating states of a consumer;  
       FIGS. 4 and 5  show advantageous embodiments of the drive of a load-holding brake valve; and  
       FIG. 6  shows an alternative advantageous embodiment.  
    
    
      In  FIG. 1 , which is a schematic diagram,  1  designates a hydraulically controlled valve, which, in this exemplary embodiment, is a load-holding brake valve. The view of this hydraulically controlled valve  1  in the form of a load-holding brake valve does not reveal any of the internal structure of the valve, since this internal structure is not essential to the invention and is known in and of itself from WO 97/32136 A1. Omitting a diagram of the internal structure is appropriate also because the parts of the hydraulically controlled valve  1  not essential to the invention could also be of a design completely different from that illustrated and described in WO 97/32136 A1. The invention is therefore independent of a specific design of the load-holding brake valve and thus completely independent of the design of the valve  1 . The only essential point is that the valve  1  can be hydraulically controlled by at least one hydraulic drive and that the valve  1  has a flow-control device  2 , by which the flow of hydraulic oil from and to a consumer can be controlled. This flow-control device  2  can be controlled by a hydraulic drive  3 . The parts of this drive  3  essential to its function include a control plunger  4 , which is part of a control piston  5 , which acts on the flow-control device  2 . If the valve  1  is a load-holding brake valve, also called a countertorque brake valve, the flow-control device  2  consists, for example, of a pilot valve and a main valve. If the valve  1  is of a different design, different parts will be present. In the case of a directional control valve according to WO 02/075162 A1, for example, the control plunger  4  acts directly on a slide piston.  
      A side view of the control piston  5  is shown. It is designed according to the invention as a stepped piston, the inventive features of which are described below. It should be mentioned beforehand, however, that a control pressure connection X is present in a housing part  6  on the left side of the valve  1 . A bore, designated here the primary control pressure chamber  7 , is provided in the housing part  6  at the control pressure connection X.  
      According to the invention, the control piston  5  has a first step  8  on the end facing the control pressure connection X; the diameter D 8  of this step is smaller than the inside diameter of the primary control pressure chamber  7  but only just enough to allow the piston to move. A control pressure P X , which is present at the control pressure connection X and which therefore acts in the primary control pressure chamber  7 , exerts a force F on the control piston  5 . This force is equal to the product of the control pressure P X  and the end surface area A 8  of the first step  8 , where the end surface area A 8  of the first step  8  is the product of half the diameter D 8  squared times a. The control pressure P X  therefore produces a force F by which the control piston  5  is pushed against a control spring  9 . The distance which the control piston  5  travels therefore depends on the spring rate of the control spring  9 .  
      According to the invention, the control piston  5  has a second step  10 , the diameter D 10  of which is larger than the diameter D 8 . The diameter D 10  is slightly smaller than the inside diameter of a bore in the housing part  6 . This bore in the housing part  6  is designated the secondary control pressure chamber  11 . The additional hydraulically active surface area A 10  of this second step  10  is a circular ring with the outer diameter D 10  and the inner diameter D 8 .  
      It is essential to the invention that the primary control pressure chamber  7  and the secondary control pressure chamber  11  are connected by a connection  12  with a throttle point  13 , which is indicated schematically in  FIG. 1 . What is not essential is whether this primary control pressure chamber  7  and the secondary control pressure chamber  11  are formed by bores in a housing part  6  or whether they are realized in some other way. An alternative exemplary embodiment will be presented further below. The only point essential to the invention is that the hydraulic drive  3  has the primary control pressure chamber  7  and the secondary control pressure chamber  11 .  
      In the following description of the function of the device, it is assumed that the system is in a state of equilibrium, in which, as a result of a certain control pressure P X , the control piston  5  has taken up a certain position. A state of equilibrium also means that the control pressure P X  is present both in the primary control pressure chamber  7  and in the secondary control pressure chamber  11 , because the pressure has become equalized through the connection  12  containing the throttle point  13 . When the control pressure P X  is now increased, the force acting on the end surface A 8  also increases, which causes the control piston  5  to move toward the right against the control spring  9 . At this moment, however, the higher control pressure P X  is present only in the primary control pressure chamber  7 . Because of the throttle point  13 , the pressure in the secondary control pressure chamber  11  cannot increase immediately. On the contrary, when the higher control pressure P X  in the primary control pressure chamber  7  causes the control piston  5  to move toward the right, the pressure in the secondary control pressure chamber  11  will fall, which opposes the movement of the control piston  5  toward the right. Only after hydraulic oil has been able to flow from the primary control pressure chamber  7  into the secondary control pressure chamber  11  through the connection  12  with the throttle point  13  will this pressure drop be compensated, and only after the arrival of additional hydraulic oil will it finally be achieved that the pressure in the secondary control pressure chamber  11  is exactly the same as the control pressure P X  also present in the primary control pressure chamber  7 . Thus a state of equilibrium is reached again, in which the control piston  5  has now taken up a new position corresponding to the higher control pressure P X .  
      During the first moment, therefore, a higher control pressure P X  acts only on the smaller end surface A 8 . Only after the pressure has equalized across the throttle point  13  does the higher control pressure P X  act also on the hydraulically active surface of the second step  10  and therefore also on the surface area A 10 , which is derived directly from the diameter D 10 . It follows from this that there is a certain delay in the movement of the control piston  5  or that this movement is damped. As a result, the task of the invention is accomplished in a surprisingly simple way, for, as a result of this damping, the valve  1  has become insensitive to internally or externally induced oscillations, without any impairment to its response sensitivity, which could not be excluded in the case of the metering valve according to WO 97/32136 A.  
      The diameter D 8  can be, for example, 14 mm; the diameter D 10  can be 20 mm. The hydraulically active surface areas A 8  and A 10  will then be 153.9 and 314.2 mm , respectively, which results in an area ratio of 1:2.04. This indicates how large the amplitude of the oscillations which can be leveled out can be.  
      The damping is similar when the control pressure P X  decreases. When the control pressure P X  is reduced, the pressure in the secondary control pressure chamber  11  can decrease slowly only as a result of the flow of hydraulic oil via the connection  12  with the throttle point  13  from the secondary control pressure chamber  11  to the primary control pressure chamber  7 .  
      There is therefore no need for the measures described in WO 97/32136 A1 to prevent the excitation of oscillations, such as the use of a nozzle and a metering valve which can be adjusted by means of an adjusting spindle. In this sense the inventive solution is extremely simple. The need to select the size of the nozzle for the specific application and to install it is also eliminated, nor is there any need for the time-consuming work of adjusting the metering valve.  
      It is advantageous to use the first step  8  of the control piston  5  in conjunction with the associated bore in the housing part  6 , which forms the primary control pressure chamber  7 , as the connection  12  containing the throttle point  13 . This is shown in  FIG. 2 . This diagram is not to drawn to scale for the sake of clarity. The primary control pressure chamber  7  has an inside diameter D 7 . The first step  8  of the control piston  5 , as already shown in  FIG. 1 , has an outside diameter D 8 . Thus a ring-shaped gap  14  is present in between, the dimensions of which are defined by the inside diameter D 7  and the outside diameter D 8 . When this ring-shaped gap  14  is used as the throttle point  13 , a remarkable advantage is obtained. Whereas a nozzle used as a throttle point  13  can change its behavior over the course of time as a result of the deposition of suspended matter, which causes a change in the throttling action, the ring-shaped gap  14  is cleaned repeatedly of deposits of suspended manner by the movement of the control piston  5  during the operation of the valve  1  ( FIG. 1 ). The throttling action thus remains constant.  
      Because the ring-shaped gap  14  is essential to the function of the device, the tolerances of the inside diameter D 7  and the outside diameter D 8  are very important. These tolerances are selected so that the ring-shaped gap  14  has a width of advantageously about 0.01-0.04 mm. To achieve this, it is possible under certain conditions to match the control piston  5  to the housing part  6  through the selection of compatible stock parts.  
       FIGS. 3   a - 3   c  show a hydraulic circuit with a consumer  20 , which, in the example illustrated here, is a double-acting cylinder with a pressure space at the bottom of the piston and another pressure space on the piston rod side. It would also be possible, however, to operate a hydraulic motor as the consumer  20  instead of the double-acting cylinder. The hydraulic circuit is shown in three different operating states, namely, the neutral position in  FIG. 3   a , the load-raising mode in  FIG. 3   b , and the load-lowering mode in  FIG. 3   c . The individual elements of the hydraulic circuit which are present are the same in all cases. The hydraulic circuit is known in and of itself and is shown here because the inventive action of the inventive hydraulically controlled valve can be described clearly on the basis of this circuit.  
      A directional control valve  21  and a load-holding brake valve  22 , which serve to control the consumer  20 , are shown in all three  FIGS. 3   a - 3   c . The load-holding brake valve  22  can be of the design described in, for example, WO 97/32136 A1, but it is equipped with a hydraulic drive  3  designed in accordance with the invention. The directional control valve  21  can be one of the types described in WO 02/075162 A1, for example, but it is also equipped with hydraulic drives  3 ′ designed in accordance with the invention.  
      The hydraulic oil can be conveyed by a pump  24 , driven by a motor  23 , between the tank  25  and the consumer  20 . The pump  24  has a first check valve  26  and a pressure-limiting valve  27  in the conventional manner. The flow of hydraulic oil is determined by the positions of the directional control valve  21  and of the load-holding brake valve  22 . A second check valve  28  is installed in the line leading to the bottom pressure space of the consumer  20 . This separate check valve  28  can be omitted if the load-holding brake valve  22  already has a check valve, which is designated in the diagram of the load-holding brake valve  22  by the reference symbol  28 ′.  
      The directional control valve  21  is controlled in the conventional manner through the actuation of its two drives  3 ′. If neither of the drives  3 ′ is actuated, that is, if a control pressure P St  is not being applied to either of them, the directional control valve  21  assumes the neutral position.  
      In the neutral position of the directional control valve  21  shown in  FIG. 3   a , the connection in the directional control valve  21  between the pump  24 , the bottom pressure space of the consumer  20 , the piston-side pressure space of the consumer, and the return flow to the tank  25  is open. This does not apply in general and is different in the case of, for example, the directional control valve according to WO 02/075162 A1. This is not important, however, with respect to the invention. For the present circuit, the only important point in terms of the correct control of the consumer  20  is that, in the neutral position, the load-holding brake valve  22  is closed, so that the consumer remains in its original position. That the load-holding brake valve  22  remains closed can be derived directly from the fact that the control pressure P X  ( FIG. 1 ) is approximately the same as the pressure in the piston rod-side pressure space of the consumer  20 , which for its own part is approximately the same as atmospheric pressure, because the connection to the tank  25  is open.  
       FIG. 3   b  shows the load-raising mode. This is reached by the actuation of one of the drives  3 ′ of the directional control valve  21  by a control pressure P St . The slide piston of the directional control valve  21  is moved in such a way that hydraulic oil can flow from the pump  24  through the directional control valve  21  to the bottom pressure space of the consumer  20  and from the piston rod-side pressure space of the consumer  20  to the tank  25 . The pump  24  therefore conveys hydraulic oil from the tank  25  to the bottom side of the consumer  20 , where the first check valve  26  and the second check valve  28  or the check valve  28 ′ are automatically actuated by the pump pressure. Because the hydraulic oil is conveyed to the bottom pressure space of the consumer  20 , hydraulic oil is simultaneously displaced from the piston rod-side pressure space of the consumer  20  and flows via the directional control valve  21  to the tank  25 .  
      The load-holding brake valve  22  has no function here. This is related to the fact that the active control pressure P X  is very low, because the hydraulic oil flows from the piston rod-side of the consumer  20  to the pressureless tank  25 , as explained in connection with the neutral position. Thus the oscillation-damping action of the drive  3  of the load-holding brake valve  22  also remains without effect.  
      If the drives  3 ′ of the directional control valve  21  are designed according to the invention, they will also produce a damping action, which is advantageous when the control pressure P St , as is often the case, is derived from the load pressure at the consumer  20  or from the pump pressure. Variations in this load or pump pressure are therefore damped in the drive  3 ′ of the directional control valve. The advantageous action of this damping occurs when, in the load-raising mode, the consumer  20  or the device operated by it encounters an obstacle which causes the load pressure to change instantaneously.  
       FIG. 3   c  shows the load-lowering mode. Here the pump  24  conveys hydraulic oil to the piston rod-side pressure space of the consumer  20 . This is achieved by the application of a control pressure P St  to the other drive  3 ′ of the directional control valve  21 . As a result, the connection in the directional control valve  21  from the pump  24  to the piston rod-side pressure space of the consumer  20  is open, and the connection from the bottom pressure space of the consumer  20  to the tank  24  is also open. The control pressure P X  acting on the load-holding brake valve  22  is now high. It is determined by the pressure generated by the pump and the pressure loss across the directional control valve  21 .  
      Because hydraulic oil is flowing to the piston rod-side space of the consumer  20 , hydraulic oil is now forced to flow from the bottom pressure space of the consumer  20  to the tank  24 . The second check valve  28 , which is parallel to the load-holding brake valve  22 , or the check valve  28 ′, however, is closed in this load situation. Hydraulic oil can therefore flow from the bottom pressure space of the consumer  20  only if the load-holding brake valve  22  is opened. This is done by the control pressure P X , the value of which is based on the proportional adjustment of the directional control valve  21  by the control pressure P St . The goal is thus achieved in the conventional manner that the hydraulic oil can leave the bottom pressure space of the consumer  20 . The quantity leaving the consumer  20  is larger than the quantity simultaneously entering the piston rod-side pressure space, because the cross section on the piston rod side is different from that on the bottom side.  
      In this operating mode, the inventive effect of the design of the drive  3  of the load-holding brake valve  22  comes into play. If the control pressure P St  is increased very quickly, the control pressure P X  also rises very quickly. The rapid increase in the control pressure P St  could cause oscillations in the consumer  20 , but this oscillation is strongly damped by the inventive design of the drive  3  of the load-holding brake valve  22 .  
      If the drives  3 ′ of the directional control valve  21  are designed as intended by the invention, the valve has a damping effect with respect to the action of the control pressure P St  on the directional control valve  21 , which has the result that, in this way, too, the tendency for oscillations to occur in the consumer  20  are eliminated. It is thus impossible for a rapid increase in the control pressure P St  to cause oscillations in the consumer  20 . Oscillations which are excited by alternating loads on the consumer  20 , however, are damped simultaneously by the drive  3  of the load-holding brake valve  22 .  
      This example shows that the inventive design of the drive  3  for the load-holding brake valve  22  can prevent oscillations during load-lowering mode. If the inventive design, which was originally intended only for use in a load-holding brake valve  22 , is also used for the hydraulic drives  3 ′ of the directional control valve  21 , additional effective damping is obtained as a result. It is therefore advantageous for the drives  3 ′ of the directional control valve  21  also to be designed in accordance with the principle of the invention.  
       FIG. 4  shows an advantageous embodiment of a drive  3 , which can be used in a load-holding brake valve  22  ( FIGS. 3   a - 3   c ).  FIG. 4  is the same as  FIG. 1  except that it also contains this advantageous embodiment. This consists in that a pressure relief check valve  30  is installed between the primary control pressure chamber  7  and the secondary control pressure chamber  11 . This makes it possible for the pressure to be released from the secondary control pressure chamber  11  to the primary control pressure chamber  7 . The pressure difference at which the pressure relief check valve  30  opens is determined by a spring  31 .  
      This pressure relief check valve  30  has the effect described below. If the control pressure P X  is reduced, as already mentioned above, the control spring  9  has the effect of moving the control piston  5  toward the left. The pressure in the secondary control pressure chamber  11  cannot fall immediately, however. The pressure drop cannot occur until the connection  12  containing the throttle point  13  becomes effective. As previously mentioned, however, the load-holding brake valve  22  does not have any effect in the load-raising state according to  FIG. 3   b . There is therefore no point in allowing the drive  3  designed in accordance with the invention to produce a damping effect in this operating situation. The pressure relief check valve  30  accomplishes this.  
       FIG. 5  is basically similar to  FIG. 4 , except that it shows the ring-shaped gap  14  instead of the connection  12  with the throttle point  13 . Here an additional advantageous embodiment is illustrated, in which a longitudinal groove  33  is cut into the cylindrical lateral surface of the first step  8  at the end facing the secondary control pressure chamber  11 . As a result of this measure, the effective length of the ring-shaped gap  14  is limited; the flow of hydraulic oil between the primary control pressure chamber  7  and the secondary control pressure chamber  11  is facilitated; and thus the action of the damping is limited. In this way, the damping action of a valve  1  can be very easily adapted to the concrete application by adjusting the length of the longitudinal groove  33  to suit the circumstances.  
       FIG. 6  shows another advantageous embodiment of a drive  3  which can be used in a load-holding brake valve  22  ( FIGS. 3   a - 3   c ). Here the pressure relief check valve  30  shown in  FIGS. 4 and 5  is integrated directly into the drive  3 . Only the parts important to the function of the inventive device are shown; the parts which, for example, serve to transmit force to the flow control device  2  to be actuated ( FIG. 1 ) and the control spring  9  ( FIG. 1 ) have been omitted.  
      What is shown is the control piston  5  with its first step  8  and its second step  10 , which, as previously explained, have the diameters D 8  and D 10 , respectively. Also shown are the primary control pressure chamber  7  and the secondary control pressure chamber  11 . In contrast to  FIG. 5 , the pressure relief check valve  30  in this exemplary embodiment is located inside the hydraulic drive  3 . In contrast to the device explained on the basis of  FIGS. 1, 4 , and  5 , the hydraulic drive  3  does not have a separate housing part  6 . Instead, the hydraulic drive  3  is located inside the housing of the valve to be controlled ( FIG. 1 ), this housing being designated by the reference number  40  in  FIG. 6 . A cover  41  can be screwed into the housing  40 , which is open toward the left. An opening is present in this cover  41 ; this opening represents the control pressure connection X, which, as also in the previous exemplary embodiments, is connected to the primary control pressure chamber  7 .  
      It is advantageous here to install an orifice  42  between the control pressure connection X and the primary control pressure chamber  7 , namely, inside the cover  41 . This orifice has the effect of limiting the flow, which means in turn that, when the control pressure P X  increases very quickly, the increase in the pressure in the primary control pressure chamber  7  is delayed. Because this delay of the pressure increase implies a damping effect, an additional advantageous measure is obtained in terms of solving the problem in question.  
      Because the inventive damping occurs by way of the throttle point  13  ( FIG. 1 ) or the ring-shaped gap  14  and because the damping by the orifice  42  is a supplemental effect, it is advantageous for the damping by the orifice  42  to be much smaller than the damping by the throttle point  13  ( FIG. 1 ) or the ring-shaped gap  14 . It has been found that an optimal effect is obtained when, for example, the dimensions of the ring-shaped gap  14  are calculated in such a way that the gap corresponds to a nozzle with a diameter of 0.1 mm, whereas the orifice  42  corresponds to a nozzle with a diameter of 0.3-0.6 mm. At a diameter ratio of 1:3-1:6, an area ratio of 1:9-1:36 is obtained. This clearly shows that the damping by the throttle point  13  ( FIG. 1 ) or by the ring-shaped gap  14  is dominant. The orifice  42  provides an additional improvement.  
      The pressure relief check valve  30  integrated into the hydraulic drive  3  is formed by a check disk  45 , which seals a seating surface  44 . This disk is pressed by the spring  31 , already shown in  FIGS. 4 and 5 , against the seating surface  44 . The check disk  45  has a central bore  46 . Inside the bore  46  is the part of the control piston  5  which forms the first step  8 . The ring-shaped gap  14  is thus limited on one side by this bore  46  and on the other side by the diameter D 8  of the first step  8  of the control piston  5 . The rules mentioned above can also be used to calculate the dimensions of the ring-shaped gap  14 . The function of this pressure relief check valve  30  has already been described above. The closed position is shown in  FIG. 6 . The pressure relief check valve  30  opens when the control pressure P X  is reduced, as already described in conjunction with  FIG. 4 . The check disk  45  moves toward the left against the force of the spring  31  and therefore rises from the seating surface  44 . Thus hydraulic oil can flow directly from the secondary control pressure chamber  11  into the primary control pressure chamber  7 .  
      As shown in  FIG. 5 , the pressure relief check valve  30  is connected in parallel to the ring-shaped gap  14  between the primary control pressure chamber  7  and the secondary control pressure chamber  11 . This is also true in the exemplary embodiment according to  FIG. 6 . The design according to  FIG. 6  results in an advantageously compact unit.  
      The invention can be applied to all types of hydraulically controlled valves I in which oscillations might occur because of the way in which the system is controlled and/or the way in which the device such as a crane or front end loader is operated by the consumer  20 .