Abstract:
A hydraulic control valve assembly ( 10 ), method and system, characterized by control valves ( 26 ) each having a variable metering orifice; a compensator ( 28 ) that controls flow of fluid from the variable metering orifice to a work port (A,B) in response to a differential in pressures acting on opposite first and second sides of the compensator, wherein a first side receives a pressure at the downstream side of the variable metering orifice; a load sense passage ( 34 ) providing a load sense pressure; and a differential pressure controller ( 40 ) having a first inlet connected to a pump supply port and a second inlet connected to the load sense passage, the controller having a first operational mode in which load sense pressure at the second inlet is supplied to an outlet of the controller and a second operational mode in which flow from the first inlet is metered to the outlet of the controller to provide a differential control output pressure, and wherein an outlet of the differential pressure controller is connected to a pump control port and/or to the second side of the pressure compensating valve of at least one of the control valves.

Description:
RELATED APPLICATION 
     This application is a national phase of International Application No. PCT/US2011/036047 filed May 11, 2011 and published in the English language which claims the benefit of U.S. Provisional Application No. 61/333,389 filed May 11, 2010, each of which is hereby incorporated herein by reference in its entirety. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to hydraulic systems. More particularly, the present invention relates to hydraulic systems for work vehicles and especially hydraulic systems that are compensated to regulate pressure differentials existing across metering orifices of control valves within the hydraulic systems. 
     BACKGROUND OF THE INVENTION 
     Hydraulic systems are employed in many circumstances to provide hydraulic power from a hydraulic power source to multiple loads. In particular, such hydraulic systems are commonly employed in a variety of work vehicles such as excavators and loader-backhoes. In such vehicles, the loads powered by the hydraulic systems may include a variety of hydraulically actuated devices such as piston-cylinder assemblies that lower, raise and rotate arms, and lower and raise buckets, as well as hydraulically-powered motors that drive tracks or wheels of the vehicles. Although the various hydraulically actuated devices typically are powered by a single source (e.g., a single pump), the rates of fluid flow to the different devices typically are independently controllable, through the use of separate control valves (typically spool valves) that are independently controlled by an operator of the work vehicle. 
     The operation of the hydraulically actuated devices depends upon the hydraulic fluid flow to those devices, which in turn depends upon the cross-sectional areas of metering orifices of the control valves between the pressure source and the hydraulically actuated devices, and also upon the pressure differentials across those metering orifices. 
     To facilitate control, hydraulic systems often are pressure compensated, that is, designed to set and maintain the pressure differentials across the metering orifices of the control valves, so that controlling of the valves by an operator only tends to vary the cross-sectional areas of the orifices of those valves but not the pressure differentials across those orifices. Such pressure compensated hydraulic systems typically include pressure compensation valves positioned between the respective control valves and the respective hydraulically actuated devices. The pressure compensation valves control the pressures existing on the downstream sides of the metering orifices to produce the desired pressure differentials across the metering orifices. 
     Such pressure-compensated hydraulic systems normally ensure that the same particular pressure differential (e.g., a pump margin pressure) occurs across each of the control valves. Nevertheless, it may be desirable in some hydraulic systems to have a lower pressure differential across selected valves to reduce the hydraulic fluid flow through those valves. For example, in the case of an excavator, it may be desirable to provide normal hydraulic fluid flow to the cylinders that control lifting or other movement of an arm or bucket of the excavator, or to accessories of the excavator such as a trenching device, yet at the same time desirable to provide reduced hydraulic fluid flow to the hydraulic motors controlling the speeds of the tracks of the excavator so that the excavator travels at reduced speeds. Therefore, there is a need in some hydraulic systems to provide a pressure differential across metering orifices in selected control valves which is less than the pressure differential across other control valves. 
     This capability of providing adjustable control of the pressure differentials across multiple control valves in an even manner is desirable in many circumstances, since it is often desirable that multiple hydraulic devices of a hydraulic system should receive precisely identical amounts of hydraulic fluid flow when an operator sets the respective control valves identically. For example, with respect to the excavator discussed above, it would be desirable that the hydraulic motors corresponding to the left and right tracks of the excavator be driven at the exact same speed assuming that the operator of the excavator set the control valves for those motors to the same level. 
     U.S. Pat. No. 6,895,852 discloses an apparatus having a valve assembly with pressure compensated valve sections. The apparatus includes an adjustable pressure reducing valve that communicates pressure from a source (e.g., a pump) to the particular compensation valves that are coupled to the control valves for which adjustable control is desired. The opposing actuation ports of the adjustable pressure reducing valve are coupled, respectively, to the pressure applied to those particular compensation valves and to the highest load pressure plus an adjustment spring pressure. Consequently, the pressure applied to the particular compensation valves exceeds that of the highest load pressure by the adjustment spring pressure, which results in reduced pressure differentials across the control valves associated with those compensation valves. Because the adjustable pressure reducing valve is in communication with each of the particular compensation valves that are coupled to the control valves for which adjustable control is desired, and because the single adjustment spring pressure determines the operation of that adjustable pressure reducing valve, an operator only needs to make a single adjustment to the single adjustment spring pressure to produce the same changes to the pressure differentials across each of the control valves for which adjustable control is desired. Also disclosed is the use of another valve that is coupled between the adjustable pressure reducing valve, the highest load pressure and the particular compensation valves of interest. The reduction in the pressure differentials produced by the adjustable pressure reducing valve can be switched on and off by alternatively coupling the particular compensation valves to the output of the adjustable pressure reducing valve and to the highest load pressure, respectively. 
     SUMMARY OF THE INVENTION 
     The present invention provides a pressure compensated hydraulic system having differential pressure control that enables fluid flow through one or more valve sections to be adjusted, as may be in many applications. One or more principles of the invention may be applied to load sense and post-pressure compensated valves. 
     In accordance with the invention, a differential pressure controller (e.g. a differential pressure control valve) senses maximum regulated pressure (e.g. load sense pressure) downstream of one or more pressure compensating valves each associated with a respective control valve (or valves) that has a variable metering orifice through which hydraulic fluid flows between an inlet port providing for connection to a pump and a respective work port providing for connection to a respective actuator (e.g. a hydraulically actuated device such as a piston-cylinder assembly, hydraulic motor, etc.). The differential pressure controller produces an output pressure that may be supplied to a pump control port to which can be connected the control port of a variable displacement pump that produces an output pressure of the pump that is a predefined amount greater than the pressure supplied to the control port of the pump. Additionally or alternatively the output pressure of the differential pressure controller may be supplied to the pressure compensating valve of at least one of the control valves. The output pressure may be equal to the sum of the maximum regulated pressure (e.g., the load sense pressure) and a setting pressure of the differential pressure controller. 
     Accordingly, the differential pressure controller may supply an output pressure that is higher than the maximum regulated or load sense pressure to either or both the pump port and a work section compensator spring chamber to change a pressure differential. Consequently, either the inlet pressure and/or pressure downstream of the work section flow output controlling area increases. In other words, the system enables the hydraulic pressure differential between the control valve inlet and the work section with the highest work port pressure and/or across one or more work section flow areas to vary flow output of the control valve. 
     Hence, according to one aspect of the invention, a hydraulic control valve assembly comprises plural control valves each having a variable metering orifice through which hydraulic fluid flows between an inlet port providing for connection to a pump and a respective work port providing for connection to a respective actuator; a compensator that controls flow of fluid from the variable metering orifice to the work port of each control valve in response to a differential in pressures acting on opposite first and second sides of the compensator, wherein the first side receives a pressure at the downstream side of the variable metering orifice; a load sense passage connected to the control valves to provide a load sense pressure corresponding to the greatest pressure amongst the work ports; and a differential pressure controller having a first inlet connected to the pump supply port and a second inlet connected to the load sense passage, the differential pressure controller having a first operational mode in which load sense pressure at the second inlet is supplied to an outlet of the differential pressure controller and a second operational mode in which flow from the first inlet is metered to the outlet of the differential pressure controller to provide a differential control output pressure at the outlet of the differential pressure controller, and wherein the outlet of the differential pressure controller is connected to (a) a pump control port to which can be connected the control port of a variable displacement pump that produces an output pressure of the pump that is a predefined amount greater than the pressure supplied to the control port of the pump, and/or (b) to the second side of the pressure compensating valve of at least one of the plural control valves. 
     Each control valve may have a respective compensator, and each compensator that is not connected to the outlet of the differential pressure controller, has the second side connected to load sense passage. 
     The differential pressure controller may be configured to provide a differential control output pressure that is greater than the load sense pressure. 
     The differential pressure controller may include a controller valve that in the second operational mode provides a pressure drop corresponding to a control force applied to the controller valve. In a particular embodiment, the control force is selected to provide a predetermined pressure difference between the differential control output pressure and the load sense pressure. 
     The control force may be provided by a control device that may be configured to provide different control forces during shifting of the differential pressure controller between its first and second operation modes. 
     The controller valve is biased to a position corresponding the first operation mode. 
     According to another aspect of the invention, a method of controlling a hydraulic system wherein plural control valves each have a variable metering orifice through which hydraulic fluid flows between an inlet port providing for connection to a pump and a respective work port providing for connection to a respective actuator, comprises the steps of using a compensator to control flow of fluid from the variable metering orifice to the work port of each control valve in response to a differential in pressures acting on opposite first and second sides of the compensator, wherein the first side receives a pressure at the downstream side of the variable metering orifice; providing a load sense pressure corresponding to the greatest pressure amongst the work ports; and using a differential pressure controller having a first inlet connected to the pump supply port and a second inlet connected to the load sense passage, the differential pressure controller having a first operational mode in which load sense pressure at the second inlet is supplied to an outlet of the differential pressure controller and a second operational mode in which flow from the first inlet is metered to the outlet of the differential pressure controller to provide a differential control output pressure at the outlet of the differential pressure controller, and wherein the outlet of the differential pressure controller is connected to (a) a pump control port to which can be connected the control port of a variable displacement pump that produces an output pressure of the pump that is a predefined amount greater than the pressure supplied to the control port of the pump, and/or (b) to the second side of the pressure compensating valve of at least one of the plural control valves. 
     According to a further aspect of the invention, a valve assembly comprises multiple working sections, each working section having a movable control spool and a compensator, an input conduit for supplying fluid to the working sections, a load sense conduit adapted to receive a pressure signal from the working section outputting a highest pressure, and a control mechanism connected to the input conduit and the load sense conduit and provide an output pressure in response to actuation of an associated input. 
     The associated input may be a proportional solenoid, and/or the output pressure may be provided to one of a pump, a pressure gain mechanism, or a compensator of at least one of the working sections. 
     A control valve assembly employing a control mechanism according to the present invention has numerous applications. For example, with reference to a mini-excavator such as those often available for rental, a device that is actuatable by an operator for either increasing or decreasing fluid flow through one or more sections of a valve assembly may allow the mini-excavator to have multiple operating modes. In one example, the mini-excavator may have a novice operating mode and an expert operating mode, where selection of the operating mode is provided by a switch within a cab of the mini-excavator. Actuation of the switch into the novice operating mode operates to slow the speed associated with each function of the mini-excavator; whereas actuation of the switch into the expert operating mode operates to increase the speed associated with each function of the mini-excavator as compared to the speed in novice mode. 
     Further features and advantages of the invention will become apparent from the following detailed description when considered in conjunction with the drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is schematic diagram showing an exemplary hydraulic system according to the present invention, where a differential pressure controller is used to provide flow boost for all work sections. 
         FIG. 2  is schematic diagram showing another exemplary hydraulic system according to the present invention, where the differential pressure controller is used to where the differential pressure controller provides flow boost and a gain adjustment mechanism for slowing response of a controlled actuator. 
         FIG. 3  is schematic diagram showing still another exemplary hydraulic system according to the present invention, where the differential pressure controller is used to boost output flow of one or more of the work sections with low differential pressure default. 
         FIG. 4  is schematic diagram showing a further exemplary hydraulic system according to the present invention, where the differential pressure controller is used to boost output flow of one or more of the work sections with high differential pressure default. 
         FIG. 5  is schematic diagram showing a still further exemplary hydraulic system according to the present invention, where the differential pressure controller is used to reduce output flow of all work sections with high differential pressure default. 
         FIG. 6  is schematic diagram showing yet another exemplary hydraulic system according to the present invention, where the differential pressure controller is used to deactivate one or more work sections. 
     
    
    
     DETAILED DESCRIPTION 
       FIG. 1  illustrates a valve assembly  10  having multiple valve sections. The valve assembly  10  of  FIG. 1  includes an inlet section  12 , an outlet section  14 , and two working sections  16  and  18 . In  FIG. 1 , the two working sections  16  and  18  are interposed between the inlet section  12  and the outlet section  14 . Although  FIG. 1  illustrates a valve assembly  10  with only two working sections located between the inlet section and the outlet section, any number of working sections may be provided. 
     The valve assembly  10  forms a portion of a hydraulic system  100 . The hydraulic system  100  also includes a variable displacement hydraulic pump  102 , a reservoir  104 , and hydraulically actuated devices (not shown) (also herein referred to as actuators), one of which is associated with each working section  16  and  18  of the valve assembly  10 . The hydraulically actuated devices may be piston-cylinder assemblies, hydraulic motors, etc. The hydraulically actuated devices may be those used to lower, raise and rotate arms, lower and raise buckets, or to power drive tracks or wheels of vehicles, in particular excavators. 
     The hydraulic pump  102  is responsive to a pressure signal at load sense port  103  for controlling a pressure at its outlet port. For example, the hydraulic pump  102  may be designed to provide a 300 psi margin pressure. In such an example, the hydraulic pump  102  is operable to maintain an outlet pressure that is 300 psi greater than the received pressure. The pump  102  adjusts its displacement so as to maintain the margin pressure based outlet pressure. 
     In other embodiments, other types of load sensing margin pressure sources may be used. For example, a fixed displacement pump may be used with a bypass valve that modulates flow bypassed back to the reservoir in response to a pressure signal, whereby the pressure supplied at the outlet of the pressure source maintains a pressure that is greater than the pressure signal by a prescribed amount. Such a load sensing margin pressure source may be used interchangeably with the herein illustrated load sensing margin pressure sources using variable displacement pumps. 
     The outlet port of the hydraulic pump  102  is in fluid communication with the inlet section  12  of the valve assembly  10 . An inlet conduit  24  of the valve assembly  10  includes an inlet port  25  preferably located in the inlet section  12 . The inlet conduit  24  extends through the inlet section  12 , through each working section  16  and  18 , and into the outlet section  14  of the valve assembly  10 . 
     Each of the working sections  16  and  18  of the valve assembly  10  includes an associated control spool  26  and an associated compensator  28 . In the embodiment illustrated in  FIG. 1 , the compensator  28  of each working section  16  and  18  is located downstream of the control spool  26  relative to the inlet conduit  24 . Thus,  FIG. 1  illustrates post-pressure compensated working sections. 
     The inlet conduit  24  provides fluid to the control spool  26  of each working section  16  and  18 . The control spools  26  are independently actuatable to move from a neutral, closed position to a position for directing hydraulic fluid toward the compensator  28  of the associated working section.  FIG. 1  schematically illustrates handles  30  mechanically linked to each of the control spools  26  for moving the control spools in response to operator inputs. Alternatively, the control spools  26  may be moved via an indirect linkage so that the operator may be positioned remote from the valve assembly. As will be appreciated by those skilled in the art, other means may be used to control movement of the spool, including electrically actuated members such as a solenoid that may be pulse width modulated in response to movement of a control member located, for example, in the cab of a vehicle. 
     In response to movement of a control spool  26  of a working section  16  and  18 , fluid flows from the inlet conduit  24  across the control spool  26  and into a metered cavity of the working section located immediately upstream of the compensator  28 . A pressure drop occurs as the fluid passes across the control spool  26  to the metered cavity. 
     The compensator  28  of each working section  16  and  18  is adapted to maintain a set pressure drop within the working section. The set pressure drop is related to a received pressure signal, commonly called a load sense signal. As illustrated in  FIG. 1 , each compensator  28  receives the load sense signal from a load sense conduit  34 . The load sense signal corresponds to the highest working pressure from the work ports of the valve assembly  10 . 
     Thus, with reference to the exemplary embodiment of  FIG. 1 , the load sense signal will equal 2000 psi, which is the pressure in the work port of working section  16  that is being supplied with fluid. Load sense check valves  35  in the working sections of the valve assembly  10  are arranged such that the highest work port pressure is provided into the load sense conduit  34 . 
     In addition to receiving the load sense signal from the load sense conduit  34 , each compensator  28  also includes a spring  36  having a preset spring force for biasing a poppet of the compensator  28  into a closed position, as is illustrated schematically in  FIG. 1 . Thus, the pressure applied by the spring  36  is added to the pressure applied by the load sense signal for biasing the compensator  28  into a closed position. The compensator  28  is opened in response to the fluid pressure in the metered cavity increasing to a value which is greater than the sum of the spring pressure and load sense signal. When the compensator  28  is opened, fluid flows past the poppet of the compensator and to a work port of the associated working section of the valve assembly  10  preferably via a load check valve  39 . In the embodiment of  FIG. 1 , both working sections  16  and  18  have identical configurations. Each working section has respective work ports A and B that provide for connection to inlet/outlet ports of a hydraulically actuated device whereby fluid can be supplied to and returned from the device. Return flow is directed to an outlet conduit  41  extends through the inlet section  12 , through each working section  16  and  18 , and into the outlet section  14  of the valve assembly  10 . The outlet conduit  41  is connected to an outlet port  43  that provides for connection to the reservoir  104  or directly to the inlet of the pump  102 . 
     The outlet section  14  of the valve assembly  10  according to the embodiment of  FIG. 1  also receives fluid from the inlet conduit  24 . The outlet section  14  further includes a differential pressure control mechanism  40 , herein also referred to as a differential pressure controller. In the embodiment of  FIG. 1 , the control mechanism  40  is operable for controlling fluid pressure to a control conduit  42  that leads to a control port  45  which connected by a line  46  to the load sense port  103  of the pump  102 . The control mechanism  40  is adapted to receive pressure from two inputs: (i) the load sense conduit  34  and (ii) the input conduit  24 . The control conduit  42  in  FIG. 1  provides a pressure signal to the load sense port of the pump  102  that is the basis for controlling the outlet pressure of the pump. More specifically, the pump  102  attempts to maintain an outlet pressure that exceeds the pressure in the control conduit  42  by the margin pressure. 
     The control mechanism  40  includes a first position, illustrated schematically in  FIG. 1 , in which the load sense conduit  34  is in communication with the control conduit  42 . When the control mechanism  40  is in the first position, there is no pressure drop (or only a negligible pressure drop) across the control mechanism  40  and, the load sense signal pressure from the load sense conduit  34  is transferred to the control conduit  42 . As a result, the pump  102  operates to provide an outlet pressure that exceeds the load sense signal pressure by the margin pressure. 
     The control mechanism  40  moves in response to a controlled input from the first position to a position in which the input conduit  24  is in communication with the control conduit  42 . The pressure drop across the control mechanism  40  may be controlled when the input conduit  24  is in communication with the control conduit  42 . For example, in one embodiment, the controlled input is provided by an input device  48  such as a proportional solenoid (or a hydraulic or pneumatic pressure source  48   a  either within the hydraulic system or separate from the hydraulic system, an adjustable spring mechanism  48   b , or a stepper motor or similar device) that is adapted to adjust the pressure drop across the control mechanism  40 , for example in the range of 0 to 300 psi. A return spring  50  may act to return the control mechanism  40  to the first position in the absence of a higher force from the proportional solenoid  48  or other input device. Additional alternatives for the controlled input may include an adjustable pressure input from a hydraulic or pneumatic pressure source either within the hydraulic system or separate from the hydraulic system, an adjustable spring mechanism, or a bi-directional pilot valve, stepper motor or similar devices that avoid the need for the return spring, or similar devices in general. The input device may be controlled by a suitable controller such as a microprocessor, programmable controller or the like, with one or more inputs, such as a selector input for enabling selection between different modes of operation of the control mechanism. The controller may have other inputs for receiving signals from one or more sensors that report system pressures, fluid flows, states, etc. to the controller. This may include end-of-stroke sensors for use in connection with one or more of the different embodiments for automatic cylinder speed reduction at the end of stroke, as discussed further below. The controller may also provide for proportional control of the controlled input for providing desired functionality. The controller may even simply be a mode selector switch. 
     During operation of the hydraulic system  100  of  FIG. 1  with the control mechanism  40  in its illustrated position (i.e., the first position corresponding to a first operational mode), pressure from the load sense conduit  34  is provided to the control conduit  42  without a pressure drop (or with a negligible pressure drop) and, the pump  102  attempts to maintain an outlet pressure equal to the load sense pressure plus the margin pressure. Depending upon the pressure drop across the control spool  26  and the compensator  28  of each working section  16  and  18 , fluid flows across the control spool and the compensator to a working port of the associated working section for actuating the associated hydraulically actuatable device. 
     For example, assume that working section  16  is actuated for providing fluid at 2000 psi to its working port B and working section  18  is actuated for providing fluid at 1000 psi to its working port B. The load sense signal pressure, i.e., the highest working port pressure, will be 2000 psi. With the control mechanism  40  in the first position, as illustrated in  FIG. 1 , the 2000 psi load sense signal pressure is provided to the control conduit  42  and to the pump  102 . The pump  102  applies its margin pressure to the pressure received from the control conduit  42  and, as a result, attempts to maintain an output pressure of 2300 psi (when the margin pressure is 300 psi). In this example, with reference to working section  16 , the pressure drop across the control spool and the compensator equals 300 psi for providing 2000 psi to working port B. 
     Now, assume that the proportional solenoid  48  is actuated and the control mechanism  40  shifts to a position for connecting the inlet conduit  24  to the control conduit  42 , this corresponding to a second operational mode. When first actuated, the proportional solenoid  48  controls the control mechanism  40  to provide a first pressure drop between the inlet conduit  24  and the controlled conduit  42 . The proportional solenoid  48  then adjusts the pressure drop across the control mechanism  40  to provide the desired pressure in the control conduit  42 . 
     For example, still assume that working section  16  is actuated for providing fluid at 2000 psi to its working port B and working section  18  is actuated for providing fluid at 1000 psi to its working port B. The proportional solenoid  48  controls the control mechanism  40  for providing the desired pressure in the control conduit  42 . In this example, assume that the desired pressure in the control conduit  42  is 2100 psi (100 psi higher than the load sense signal pressure). Thus, when initially actuated, the proportional solenoid  48  controls the control mechanism  40  to provide a 200 psi pressure drop (2300 psi pump outlet pressure to the 2100 psi control conduit  42  pressure). In response to the 2100 psi pump control conduit pressure, the pump  102  applies its margin pressure to attempt to maintain an outlet pressure of 2400 psi. As the pump outlet pressure increases to 2400 psi, the proportional solenoid  48  adjusts to increase the pressure drop across the control mechanism  40  to 300 psi for maintaining the 2100 psi pressure in the control conduit  42 . In response to the pump outlet pressure increasing to 2400 psi, the pressure drop across the control spool  26  of each working section  16  and  18  increases (in this example, the pressure drop increases by 100 psi as the pump outlet pressure increases from 2400 psi to 2300 psi). As a result of the increase pressure drop across the control spool  26  of each working section  16  and  18 , flow to the associated hydraulically actuated devices increases and thus, the actuation speed of the associated hydraulically actuated devices increases. 
     As will be appreciated, the hydraulic system shown in  FIG. 1  has particular application to a mini-excavator such as those often available for rental. The differential pressure controller can be energized to allow for increased fluid flow and de-energized for decreased flow. In one example, the mini-excavator may have a novice operating mode and an expert operating mode, where selection of the operating mode is provided by a switch within a cab of the mini-excavator. Actuation of the switch into the novice operating mode operates to slow the speed associated with each function of the mini-excavator; whereas actuation of the switch into the expert operating mode operates to increase the speed associated with each function of the mini-excavator as compared to the speed in novice mode. 
     Before leaving  FIG. 1 , it is noted that hydraulic system  100 /valve assembly  10  may be provided with other operational components that are typically employed in similar types of valves as known to those skilled in the art. For example, the valve assembly may include a load sense pressure relief valve  83  that dumps hydraulic fluid to the reservoir  104  if the load sense pressure exceeds a prescribed amount. Associated with the relief valve  83  is an orifice  85  that limits flow to the load sense relief valve and provides dampening. The relief valve may be conveniently located in the inlet section  12 . The valve assembly  10  may also be provided in a conventional manner with a bleed orifice for decaying load sense pressure to reservoir pressure when the work sections are deactivated to let the system go to a low pressure standby mode, as is known in the art. 
     The valve assembly  10   a  of  FIG. 2  is the same as that of  FIG. 1  except for a pressure gain mechanism  60  provided in conduit  42  for receiving the outlet pressure of the differential pressure controller  40 . The pressure gain mechanism  60  is a two-position mechanism having a first flow orifice  62  associated with a first position and a second flow orifice  64  associated with a second position. The second flow orifice  64  is sized larger than the first flow orifice  62  and therefore, provides a lower pressure drop than the first flow orifice. The pressure gain mechanism  60  is biased into a first position by a return spring  66 . The pressure gain mechanism  60  is adapted to shift from the first position to a second position in response to an increase in pressure in conduit  42 . The pressure gain mechanism  60 , when located in its first position, acts to restrict flow from conduit  42  so as to slow the responsiveness of the pump  102 . In response to an input to the control mechanism  40  by the proportional solenoid  48 , the gain mechanism  60  moves from the first to the second position whereby the gain mechanism is less restrictive to flow in the conduit  42  so as to quicken the responsiveness of pump  102  to change pressure. In this (expert) mode, both the response to build pressure and work section flow output/actuator speed increase. In the other (novice) mode, the response will be less than in the expert mode. It is noted that in another embodiment the gain control mechanism can be shifted between its two states other than by means of the output of the differential pressure controller, such as by means of a programmable processor or other controller, or simply by a mode selector switch. 
     In the  FIG. 2  embodiment, pressure gain control and differential pressure control work together, for example, to assist a novice operator with more forgiving operation. For the novice, differential pressure control will limit the machine&#39;s maximum function speed while pressure gain control cushions the fast reaction of machine controls. As a result, a mini-excavator can be suitable for use by either a novice or expert operator. 
     Another embodiment is shown in  FIG. 3  which is the same as the  FIG. 1  embodiment except as noted below or is otherwise evident from the figure. In the  FIG. 3  embodiment, the compensator  28  of working section  16  receives the load sense signal from load sense line  24 , while the compensator  28  of working section  18  receives the pressure signal output from the control mechanism  40  to the control conduit  42 . In response to the proportional solenoid  48  (or other input to the control mechanism  40 ) being actuated to modify the pressure provided to the pump, the pressure drop within working section  16  increases, while the pressure drop in working section  18  remains constant. This design provides priority to working section  16  having the greater pressure drop. As a result of the increased pressure drop, more flow is provided to the hydraulically actuated device associated with working section  16 . As a result, the actuation speed of the hydraulically actuated device associated with working section  16  increases. 
     Although  FIG. 3  illustrates only one working section  18  receiving the output from the control mechanism  40 , any number of working sections may receive this output. The configuration of  FIG. 3  may be used, in one example, in a mini-excavator with one working section associated with the swing and another working section associated with the boom. As a result, when the proportional solenoid  48  is actuated to boost the pressure provided to the control conduit  42 , the pressure drop associated with the working section associated with the swing may be increased so that the swing function receives increased flow for increasing the speed of actuation. At the same time, the pressure drop associated with the working section associated with the boom remains constant whereby the boom function acts in the same manner as it did prior to actuation of the proportional solenoid  48  (or other input). 
       FIG. 4  illustrates a further embodiment of a valve assembly  10   c  constructed in accordance with the present invention. The valve assembly  10   c  is the same as  FIG. 3  except as noted below or is otherwise evident from the figure. In  FIG. 4 , the output of the control mechanism  40  is not connected to the pump  102 . Instead, the control port  103  of the pump  102  receives the load sense signal via a load sense port  61  and attempts to maintain an outlet pressure based on its established margin pressure and the load sense signal pressure. The output of the control mechanism  40  is provided to the compensator  28  of working section  18 . As a result, the pressure drop associated with working section  18  is decreased and less fluid flows through working section  18  to its associated hydraulically actuated device. The actuation speed of the associated hydraulically actuated device, in turn, slows due to the decreased flow. 
     Although  FIG. 4  illustrates only one working section receiving the boosted (higher pressure) output from the control mechanism  40 , any number of working sections may receive this boosted output and any number may not receive the boosted output. 
     For example, in the valve assembly  10   d  shown in  FIG. 5 , the boosted output of the control mechanism  40  is provided to both working sections  16  and  18  of the valve assembly  10   e . The  FIG. 5  embodiment is otherwise the same as the  FIG. 4  embodiment. 
     The valve assembly  10   e  of  FIG. 6  is similar to that of  FIG. 4 ; however, the control mechanism  40  of  FIG. 6  is actuatable for deactivating one of the working sections. Although  FIG. 6  illustrates only one working section  18  being deactivated by the boosted output from the control mechanism  40 , any number of the working section may receive this boosted output. 
     As shown in  FIG. 6 , the boosted output of the control mechanism  40 , when activated, is provided to the compensator  28 . The boosted output has a pressure sufficient to maintain the compensator  28  in a working section  18  in a closed position. As a result, no fluid flows through the working section  18  and, the hydraulically actuated device is deactivated. The configuration of  FIG. 6  may be used, for example, to provide a mode of operation in which one or more functions are deactivated so as to prevent accidental actuation. Such mode may be useful during, for example, towing a vehicle. 
     The control mechanism  40  can also be used to provide a programmed damping mode by providing automatic cylinder speed ramp-down near the end of stroke, thereby extending the component and overall machine life. That is, cylinder movement can be gradually or quickly slowed near the end of stroke to prevent hard impact. This can be effected by varying the control input to the control mechanism  40  such as by means of a proportional control. In the  FIGS. 1-3  embodiments, the control mechanism  40  can be converted from an energized to a de-energized state near the end of stroke of a piston-cylinder assembly, or from a de-energized state to an energized state in the  FIGS. 4 and 5  embodiments. Proportional control of the control mechanism  40  can be used to change other dynamics of the valve assemblies  10 ,  10   a , . . . and/or the associated systems, as may be desired. 
     Although the invention has been shown and described with respect to a certain preferred embodiment or embodiments, it is obvious that equivalent alterations and modifications will occur to others skilled in the art upon the reading and understanding of this specification and the annexed drawings. In particular regard to the various functions performed by the above described elements (components, assemblies, devices, compositions, etc.), the terms (including a reference to a “means”) used to describe such elements are intended to correspond, unless otherwise indicated, to any element which performs the specified function of the described element (i.e., that is functionally equivalent), even though not structurally equivalent to the disclosed structure which performs the function in the herein illustrated exemplary embodiment or embodiments of the invention. In addition, while a particular feature of the invention may have been described above with respect to only one or more of several illustrated embodiments, such feature may be combined with one or more other features of the other embodiments, as may be desired and advantageous for any given or particular application.