Patent Description:
Each main flow control valve can include a main flow control spool that is operated in response to the input command to control fluid flow and pressure to one or more associated hydraulic fluid receiving devices of the machinery. The hydraulic fluid receiving devices can include one or more hydraulic storage devices such as tanks or accumulators, hydraulic linear or rotary actuators, other hydraulic valves or subsystems, and/or any other devices that receive hydraulic fluid.

In examples, the machinery can include a plurality of main flow control valves for supplying and/or operating different hydraulic actuators in a hydraulic system of the machinery. A main flow control valve and its associated controls can be incorporated into a valve housing, and each such valve assembly is referred to as a worksection. Worksections of the same or different configuration may be combined, for example in a side-by-side arrangement. A worksection combined with other sections (for example, other worksections, an inlet section, and an outlet section) can be referred to as an assembly of valve sections.

A worksection operates by controlling the cross-sectional area of a main flow control valve variable area orifice. In examples, the main control valve variable orifice is located in a fluid flow path extending between an inlet passage and an outlet passage, or workport. The inlet passage may be connected directly or indirectly to a source of fluid flow and pressure, and the outlet passage may be connected directly or indirectly to one or more of the hydraulic actuators. The flow through a given main valve orifice area is dependent upon the pressure drop across the orifice. If the pressure drop across the orifice changes, the fluid flow through the orifice can vary for the same size of orifice.

Further, when a single actuator is actuated by an operator, the amount of flow going to the actuator can be proportional to a command provided by the operator, e.g., through a joystick or a lever. However, when multiple actuators are actuated simultaneously, the flow provide by a source of fluid (e.g., a pump) is divided between the multiple actuators. The actuator subjected to the smallest load might receive the highest amount of flow as fluid seeks the path of least resistance. However, this can lead to unpredictable, undesirable performance as the amount of fluid provided to a particular actuator is no longer proportional to the command provided by the operator.

It may thus be desirable to have a valve where proportionality between the command provided by the operator and the amount of flow to a particular actuator is maintained even when multiple actuators are actuated at the same time. A prior art directional control valve assembly is disclosed in <CIT>. The prior art valve assembly disclosed has a pressure compensation valve in which there are provided a directional control valve in which a main spool is slidably inserted in a spool bore that is formed with a pump port, a first and a second load pressure detecting port, a first and a second actuator port, and a first and a second tank port.

The present invention is about a pressure compensator valve, and a worksection and a valve assembly comprising such a pressure compensator valve, and is set forth in the appended claims.

In a first example implementation, the present invention describes a pressure compensator valve according to claim <NUM>.

In a second example implementation, the present invention describes a worksection of a valve assembly according to claim <NUM>.

In a third example implementation, the present disclosure describes a valve assembly according to claim <NUM>.

In addition to the illustrative aspects, implementations, and features described above, further aspects, implementations, and features will become apparent by reference to the figures and the following detailed description.

The novel features believed characteristic of the illustrative examples are set forth in the appended claims. The illustrative examples, however, as well as a preferred mode of use, further objectives and descriptions thereof, will best be understood by reference to the following detailed description of an illustrative example of the present disclosure when read in conjunction with the accompanying Figures.

In certain applications, hydraulic fluid flow in a hydraulic machine can be controlled using hydraulic sectional control valves. A sectional control valve or valve assembly can include a plurality of separate cast and machined metal valve worksections. Each worksection may include internal fluid passages, external ports, and valve bores with valve members slidably disposed within each valve bore. The valve bores may include a main control valve spool bore in which a main directional control valve spool is slidably disposed. Each worksection may be configured to control flow of fluid to and from a hydraulic actuator of the hydraulic machine.

A pressure compensated worksection is a worksection that includes a pressure compensator valve arranged to maintain a substantially predetermined pressure drop across the main control valve variable orifice under normal operating flow conditions independently of the inlet or outlet pressure. By maintaining this substantially constant pressure drop across the orifice, a constant and repeatable flow rate through the orifice is achieved for any orifice area that is determined by the input command. Pressure compensated worksections such as those described above may, for example, be a pre-compensated working section including a pressure compensator valve located prior to (or upstream of) the main valve variable orifice.

Worksections can also include load-sense passages. The load-sense passages can be configured to transmit a pressure feedback signal from an outlet passage, which indicates the fluid pressure required by the fluid flow receiving device controlled by the valve. The load-sense passage can be operably connected to a load sensing variable displacement hydraulic pump or other load sensing source of pressure and flow to provide a feedback signal to the source. Further, an outlet passage's pressure feedback signal may be connected to a pressure compensator spool of the pressure compensator valve. The pressure compensator valve then operates to maintain the predetermined pressure drop by sensing the downstream (or outlet passage) and upstream pressures across the variable orifice.

In some cases, multiple worksections are actuated at the same time (when main spools of the respective worksections are actuated simultaneously), and the total amount of fluid requested by all worksections can exceed the flow capacity of the source of fluid (e.g., the pump). In these cases, the pressure compensator valve of each section can operate to maintain pressure drop locally (for the respective section including the pressure compensator) across the main spool of the worksection without being "aware" of loads experienced by the other worksections. As such, the worksection controlling an actuator having the lowest load can allow a larger amount of fluid flow to the worksection compared to the other worksections as fluid seeks a path of least resistance. This may cause the actuator controlled by such worksection to move faster than the other actuators controlled by the other worksections.

In other words, the amount of fluid flow and speed of a given actuator may no longer be proportional to a position of the main spool within the worksection. Rather, the amount of fluid flow to a worksection depends on the loads that the other actuators are subjected to. Such loss of proportionality and predictability may be undesirable.

It may thus be desirable to have a valve where the pressure compensator valve of each worksection responds, not only to local load to which the actuator associated with the worksection is subjected to, but also to the highest load-induced pressure amongst all actuators. This way, the pressure compensator valves can respond together such that when multiple worksections are actuated simultaneously, the amount of fluid flow to each actuator can be reduced proportionally when the source of fluid (e.g., the pump) cannot provide the amount of fluid flow collectively requested by all the actuators. This configuration may be referred to as "flow sharing" where the actuators proportionally share the flow from the source regardless of the loads to which the different actuators are subjected to.

Disclosed herein are systems, valve sections, and valve assemblies that achieve flow sharing between multiple worksections of a valve. Particularly, the compensator spool of a pressure compensator is subjected to at least four pressure signals: (i) a first signal representing pressure level of regulated fluid flow provided from the pressure compensator valve to the main control spool of the worksection, (ii) a second signal representing local load (i.e., load-sense signal indicative of load-induced pressure generated by the load to which the actuator controlled by the worksection is subjected to), (iii) a third signal representing the highest load-induced pressure generated by the highest load amongst all actuators operating at the same time, and (iv) a fourth signal representing inlet pressure of the source of fluid (e.g., the pump). As such, each compensator spool responds, not only to local pressure condition of the particular worksection, but to "global" pressure conditions of other worksections as well. This leads to proportional variation of fluid and sharing of fluid flow provided by the source among all worksections. The term "local" is used herein to indicate that a parameter pertains to a particular worksection in the valve assembly, whereas the term "global" is used to indicate the highest value of the parameter among all worksections of the valve assembly.

Additionally, disclosed herein is a configuration of at least one pressure compensator of a particular worksection where the pressure compensator spool is further subjected to a biasing force of a compensator spring in addition of four fluid forces resulting from the four pressure signals mentioned above. Such biasing force can increase the amount of fluid flow of the particular worksection compared to the other worksections. Further, the compensator spring can enhance compensator efficiency of the particular worksection by reducing the effects of flow forces acting on the pressure compensator spool. Additionally, the biasing force of the compensator spring can be adjusted by an adjustment mechanism to adjust the flow division between the multiple worksection, i.e., give one worksection flow priority or equalize flow between multiple worksections actuated at the same time.

<FIG> illustrates a valve assembly <NUM>, and <FIG> illustrates a schematic of a hydraulic system <NUM> that includes the valve assembly <NUM>, in accordance with an example implementation. <FIG> and <FIG> are described together.

The valve assembly <NUM> has an inlet section <NUM>, a first worksection <NUM>, a second worksection <NUM>, a third worksection <NUM>, a fourth worksection <NUM>, a fifth worksection <NUM>, and an outlet section <NUM>. The illustrated valve assembly <NUM> and the hydraulic system <NUM> are provided for illustration purposes, and in other examples, more or fewer worksections can be used.

The inlet section <NUM>, the worksections, <NUM>, <NUM>, <NUM>, <NUM>, and <NUM>, and the outlet section <NUM> can be coupled together by fasteners (e.g., bolts screws, clamps, tie rods, etc.) to provide an assembly of valve sections. For example, the worksections <NUM>-<NUM> can be positioned adjacent one another between the inlet section <NUM> and the outlet section <NUM> of the valve assembly <NUM>. The outlet section <NUM> can receive fluid from any of the inlet section <NUM> and/or the worksections <NUM>-<NUM>.

Each of the worksections <NUM>-<NUM> has a housing or worksection body. For instance, the worksection <NUM> has a worksection body <NUM>, the worksection <NUM> has a worksection body <NUM>, the worksection <NUM> has a worksection body <NUM>, the worksection <NUM> has a worksection body <NUM>, and the worksection <NUM> has a worksection body <NUM>. The worksection bodies <NUM>, <NUM>, <NUM>, <NUM>, and <NUM> are shown schematically in <FIG> as envelope border.

The hydraulic system <NUM> shown in <FIG> can include a source <NUM> of fluid. The source <NUM> of fluid can be a pump (e.g., fixed displacement, variable displacement pump, a load-sense variable displacement pump, etc.), an accumulator, etc. The source <NUM> can be fluidly coupled to an inlet port <NUM> disposed in the inlet section <NUM> of the valve assembly <NUM> such that output fluid flow from the source <NUM> is provided to the inlet port <NUM>. The output fluid flow of the source <NUM> is then provided to the valve sections of the valve assembly <NUM> via inlet flow passage <NUM> shown in <FIG>.

In the examples where the source <NUM> comprises a pump, such pump can receive fluid from a fluid tank or reservoir <NUM>, and the pump then provides fluid flow to the valve assembly <NUM>. The reservoir <NUM> can be fluidly coupled to a reservoir port <NUM> also disposed in the inlet section <NUM> of the valve assembly <NUM>. During operation of the valve assembly <NUM>, fluid returns to the reservoir <NUM> from the valve sections of the valve assembly <NUM> via a return flow passage <NUM> and through the reservoir port <NUM>.

Referring to <FIG>, each of the worksections <NUM>-<NUM> can include a main control valve. For example, the worksection <NUM> includes main control valve <NUM>, the worksection <NUM> includes main control valve <NUM>, the worksection <NUM> includes main control valve <NUM>, the worksection <NUM> includes main control valve <NUM>, and the worksection <NUM> includes main control valve <NUM>. In examples, the main control valves <NUM>-<NUM> can be configured as three position-four way valves as depicted schematically in <FIG>.

The worksections <NUM>-<NUM> also include workports that are configured to be fluidly coupled to ports of respective hydraulic actuators (e.g., hydraulic cylinders or hydraulic motors). For example, the worksection <NUM> includes a first workport <NUM> (C1 port) and a second workport <NUM> (C2 port); the worksection <NUM> includes a first workport <NUM> (C1 port) and a second workport <NUM> (C2 port); the worksection <NUM> includes a first workport <NUM> (C1 port) and a second workport <NUM> (C2 port); the worksection <NUM> includes a first workport <NUM> (C1 port) and a second workport <NUM> (C2 port); and the worksection <NUM> includes a first workport <NUM> (C1 port) and a second workport <NUM> (C2 port). Each of the main control valves <NUM>-<NUM> is configured to control supply fluid flow from the inlet flow passage <NUM> to the respective workports C1 and C2 of the worksection and control return fluid flow from the workports to the return flow passage <NUM>.

Each of the main control valves <NUM>-<NUM> includes a respective main control spool. The main control spool can be configured to be biased to a neutral or centered position by springs. For example, the main control valve <NUM> can have springs <NUM>, <NUM> that bias its main control spool (e.g., the main control spool <NUM> described below) to a neutral position at which the main control spool can block fluid flow from the source to the workports <NUM>, <NUM>. The term "block" is used throughout herein to indicate substantially preventing fluid flow except for minimal or leakage flow of drops per minute, for example.

The main control spool can be actuated in either direction from the neutral position via various types of mechanisms. As an example for illustration, the main control spool of the worksection <NUM> can be controlled by pilot valves <NUM>, <NUM> that are solenoid-operated and can be used to actuate or move the spool in a spool bore disposed with the worksection <NUM>. However, other configurations of actuation mechanisms (e.g., manual, pneumatic, etc.) can be used.

In the example implementation shown in <FIG>, the pilot valves <NUM>, <NUM> are configured as pressure reducing valves that receive pressurized fluid at a reduced pressure level compared to inlet pressure received from the source <NUM>, then generate a pilot fluid signal that is proportional to a magnitude of an electric command provided thereto. Particularly, the valve assembly <NUM> can include a pressure reducing valve <NUM> disposed in the inlet section <NUM>. The pressure reducing valve <NUM> receives fluid from the source <NUM> via inlet pilot fluid passage <NUM> formed in the inlet section <NUM> and fluid filter <NUM>. The pressure reducing valve <NUM> reduces the pressure level of the fluid received from the inlet pilot fluid passage <NUM>, then provides a pilot source signal to a pilot fluid passage <NUM> that traverses the worksections <NUM>-<NUM>.

The pilot valves <NUM>, <NUM> (and the corresponding pilot valves of the other worksections) are fluidly coupled to the pilot fluid passage <NUM> and thus have access to the pilot source signal. If an electric command is provided to one of the pilot valves <NUM>, <NUM>, the pilot valve opens to provide a pilot signal having a pressure level that is reduced relative to the pressure level of the pilot source signal to one side of the main control spool of the worksection <NUM>. Responsively, the main control spool of the worksection <NUM> can move axially within the worksection body <NUM> to provide fluid flow from the inlet flow passage <NUM> to one of the workports <NUM>, <NUM>.

The valve assembly <NUM> is configured to be a load-sense (LS) valve. Particularly, the valve assembly <NUM> has a shuttle valve system that resolves the highest load or the highest load-induced pressure level indicative of the highest load to which the actuators controlled by the valve assembly <NUM> are subjected. The shuttle valve system then provides the load-sense (LS) signal indicative of the highest load to other components of the valve assembly <NUM> and the hydraulic system <NUM>.

In the example implementation of <FIG>, each of the worksections <NUM>-<NUM> can have a respective shuttle valve. For instance, the worksection <NUM> has a shuttle valve <NUM>, the worksection <NUM> has a shuttle valve <NUM>, worksection <NUM> has a shuttle valve <NUM>, worksection <NUM> has a shuttle valve <NUM>, and the worksection <NUM> has a shuttle valve <NUM>. In examples, each of the shuttle valves <NUM>-<NUM> can be configured to have two inlet ports and one outlet port. The shuttle valve can receive pressure signals at both inlet ports, and the pressure signal having the higher pressure level is communicated to the outlet port.

For instance, the shuttle valve <NUM> of the worksection <NUM> has a first inlet port subjected to a pressure signal indicative of the higher of the pressure levels between the workports <NUM>, <NUM> and a second inlet port that is blocked. Thus, the shuttle valve <NUM> provides a signal having the higher of the pressure levels between the workports <NUM>, <NUM> to the outlet port of the shuttle valve <NUM>. The outlet port of the shuttle valve <NUM> is in turn fluidly coupled to an inlet port of the shuttle valve <NUM> of the worksection <NUM>. The other inlet port of the shuttle valve <NUM> is subjected to a pressure signal indicative of the higher of the pressure levels between the workports <NUM>, <NUM>. Thus, a signal having the higher pressure level between the workports <NUM>, <NUM>, <NUM>, <NUM> is communicated to the outlet port of the shuttle valve <NUM>.

Similarly, the outlet port of the shuttle valve <NUM> is fluidly coupled to an inlet port of the shuttle valve <NUM> of the worksection <NUM>. The other inlet port of the shuttle valve <NUM> is subjected to a pressure signal indicative of the higher of the pressure levels between the workports <NUM>, <NUM>. Thus, a signal having the higher pressure level between the workports <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM> is communicated to the outlet port of the shuttle valve <NUM>.

The outlet port of the shuttle valve <NUM> is in turn fluidly coupled to an inlet port of the shuttle valve <NUM> of the worksection <NUM>. The other inlet port of the shuttle valve <NUM> is subjected to a pressure signal indicative of the higher of the pressure levels between the workports <NUM>, <NUM>. Thus, a signal having the higher pressure level between the workports <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM> is communicated to the outlet port of the shuttle valve <NUM>.

Similarly, the outlet port of the shuttle valve <NUM> is fluidly coupled to an inlet port of the shuttle valve <NUM> of the worksection <NUM>. The other inlet port of the shuttle valve <NUM> is subjected to a pressure signal indicative of the higher of the pressure levels between the workports <NUM>, <NUM>. Thus, a signal having the higher pressure level between the workports <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM>, <NUM> is communicated to the outlet port of the shuttle valve <NUM>. The outlet port of the shuttle valve <NUM> is fluidly coupled to a load-sense (LS) passage <NUM>. With this configuration, the LS passage <NUM> receives and transmits a pressure signal that represents a global LS pressure signal indicative of the highest pressure level among all workports.

The valve assembly <NUM>, and particularly the inlet section <NUM>, includes a restrictor check device <NUM> having a check valve and an orifice. The check valve allows fluid flow from the LS passage <NUM> to a load-sense relief valve <NUM> (and e.g., a bypass compensator valve <NUM> described below) while preventing fluid flow in the other direction. The orifice of the restrictor check device <NUM> operates as a dampening orifice to preclude or reduce instabilities that might result from pressure fluctuations in the LS passage <NUM>. The LS relief valve <NUM> prevents pressure level of the LS pressure signal in the LS passage <NUM> to exceed a threshold pressure level. If such threshold pressure level is exceeded, the LS relief valve <NUM> opens, thereby providing a flow path to the return flow passage <NUM> and the reservoir <NUM> to relieve the pressure signal in the LS passage <NUM>.

The valve assembly <NUM> further includes a flow-limiting orifice <NUM>. The flow-limiting orifice <NUM> similarly operates as a dampening orifice but also restricts the amount of fluid flowing to the LS relief valve <NUM> to preclude saturating the LS relief valve <NUM>. Particularly, if pressure level in the LS passage <NUM> exceeds the pressure setting or threshold pressure level of the LS relief valve <NUM> causing the LS relief valve <NUM> to open to the reservoir <NUM>, a pressure drop can be generated across the flow-limiting orifice <NUM>. The size of the flow-limiting orifice <NUM> then controls the amount of fluid flow thereacross (e.g., to about <NUM>,<NUM> liter per minute (lpm), i.e. <NUM>,<NUM> gallons per minute (GPM)) so as to enable the LS relief valve <NUM> to relieve the LS pressure signal and maintain the pressure level in the LS passage <NUM> at the threshold pressure level rather than being saturated.

In addition to the LS pressure signal being provided to the LS relief valve <NUM>, it is also provided in parallel to a valve load-sense passage <NUM>. The valve LS passage <NUM> is a "global" LS passage that traverses the worksection <NUM>-<NUM>. With this configuration, all the worksections <NUM>-<NUM> have access to the global LS pressure signal in the valve LS passage <NUM>, which indicates the highest pressure level or load amongst all workports <NUM>-<NUM>.

In examples, the valve assembly <NUM> can include a bypass compensator valve <NUM>. The bypass compensator valve <NUM> is configured to ensure that there is a margin pressure or differential pressure, e.g., a delta pressure of about <NUM> bar (<NUM> pounds per square inch (psi)), between pressure of fluid provided by the source <NUM> and the pressure level of the LS pressure signal in the LS passage <NUM>.

The bypass compensator valve <NUM> can have a movable element (e.g., a piston or spool) that is subjected to the LS pressure signal and pressure level in the inlet flow passage <NUM> (via the inlet pilot fluid passage <NUM>) and a biasing force of a spring, for example. The bypass compensator valve <NUM> also has an outlet port that is fluidly coupled to the return flow passage <NUM>. The movable element moves to an equilibrium position that is determined based on the pressure levels of the LS pressure signal and the inlet flow passage <NUM> and the biasing force so as to maintain a margin pressure setting or differential between the pressure level in the inlet flow passage <NUM> and the LS pressure signal. The bypass compensator valve <NUM> exhausts enough fluid flow to the return flow passage <NUM> to create the margin pressure setting. As an example for illustration, if pressure level of the LS pressure signal in the LS passage <NUM> is <NUM> bar (<NUM> psi), then the movable element (e.g., a spool or piston) within the bypass compensator valve <NUM> moves so as to cause the pressure level in the inlet flow passage <NUM> to be about <NUM> bar (<NUM> psi).

In a flow over-demand situation where the amount of fluid flow requested by all the worksections <NUM>-<NUM> exceeds a maximum flow capacity of the source <NUM>, the bypass compensator valve <NUM> might shut or close off to allow all the fluid provided from the source <NUM> to bypass the bypass compensator valve <NUM> and flow to the worksections <NUM>-<NUM>. In this case, it may take a smaller pressure differential (e.g., <NUM> bar (<NUM> psi)) rather than the pressure margin setting (e.g., <NUM> bar (<NUM> psi)) to push the maximum fluid flow capacity from the source <NUM> to the inlet flow passage <NUM>.

In another example implementation, the source <NUM> can be configured as a load-sensing variable displacement pump that receives the LS pressure signal in the LS passage <NUM>. Such pump can responsively provide enough fluid flow to maintain the margin pressure setting between its output flow provided to the inlet flow passage <NUM> and the LS pressure signal. In this example, the bypass compensator valve <NUM> might not be used in the hydraulic system <NUM>.

Further, the valve assembly <NUM> is configured such that the worksections <NUM>-<NUM> can have respective pressure compensator valves. For example, the worksection <NUM> has pressure compensator valve <NUM>, the worksection <NUM> has pressure compensator valve <NUM>, the worksection <NUM> has pressure compensator valve <NUM>, the worksection <NUM> has pressure compensator valve <NUM>, and the worksection <NUM> has pressure compensator valve <NUM>.

The pressure compensator valves <NUM>-<NUM> are disposed downstream from the inlet flow passage <NUM> and upstream from the main control valves <NUM>-<NUM>, respectively. With this configuration, the pressure compensator valves <NUM>-<NUM> can be referred to as a pre-pressure compensator valve. The pressure compensator valves <NUM>-<NUM> are configured to control supply pressure and regulate supply flow from the inlet flow passage <NUM> to the main control valves <NUM>-<NUM> so as to maintain a predetermined pressure drop across a variable metering orifice formed when a respective main control spool of a respective worksection is actuated. Maintaining a predetermined pressure drop across the variable metering orifice can allow for proportionality between the amount of flow provided to the actuator and the command signal to the pilot valve that actuates the spool, regardless of the load on the actuator and regardless of pressure level of fluid in the inlet flow passage <NUM>.

The pressure compensator valves <NUM>-<NUM> can be configured similarly or differently based on desired characteristics of a respective worksection and associated actuator. For example, in the implementation shown in <FIG>, the pressure compensator valves <NUM>-<NUM> are configured to be flow sharing compensators as described in more detail below. In particular, each of the pressure compensator valves <NUM>-<NUM> is subjected to four pressure signals and responds to a force balance based on the four pressure signals to allow for flow sharing when multiple actuators are actuated at the same time. On the other hand, the pressure compensator valve <NUM> is subjected to two pressure signals, rather than four and may thus operate differently.

Further, at least one of the pressure compensator valves <NUM>-<NUM> has a compensator spring configured to apply a biasing force on a compensator spool of the pressure compensator valve in addition to the four fluid forces applied respectively by the four pressure signals. The pressure compensator valve <NUM> has a compensator spring <NUM>. As described below, the compensator spring <NUM> enables the worksection <NUM> associated with the pressure compensator valve <NUM> to have: (i) a higher flow capacity, (ii) enhanced compensator efficiency, and (iii) priority flow compared to other worksections or equalize flow between the worksections regardless of the respective actuator loads.

For example, referring to the pressure compensator valve <NUM> of the worksection <NUM>, the pressure compensator valve <NUM> has access, or is subjected to, to four pressure signals: (i) a first signal <NUM> representing pressure level (P1) of regulated fluid flow provided from the pressure compensator valve <NUM> to the main control valve <NUM> of the worksection <NUM>, (ii) a second signal <NUM> representing local LS pressure signal having the higher pressure level (P2) between the workports <NUM>, <NUM>, (iii) a third signal <NUM> representing the highest LS pressure level (P3) transmitted via the valve LS passage <NUM>, and (iv) a fourth signal <NUM> representing inlet pressure level (P4) of fluid supplied from the source <NUM> through the inlet flow passage <NUM>.

Conventional pre-pressure compensator valves might be configured to be subjected to a local LS signal (e.g., the signal <NUM>) of the associated worksection without being subjected to the signal <NUM> representing the highest LS pressure level (P3) among all worksections. Thus, conventional pre-pressure compensator valves might not be "aware" of a loading condition in other worksections and associated actuators. Thus, in conventional valves, assuming two worksections are actuated causing a flow over-demand on the source <NUM>, if a first worksection is subjected to a lower load compared to a second worksection, more flow from the source <NUM> would flow through the first worksection compared to the second worksection.

In contrast, the pressure compensator valve <NUM> is subjected to the signal <NUM> (P3) as well as the signal <NUM> (P2). Thus, if another worksection (e.g., the worksection <NUM>) has a higher LS pressure signal, such higher LS pressure signal is communicated to the pressure compensator valve <NUM> as the signal <NUM>, and it causes the pressure compensator valve <NUM> to restrict fluid flow through the pressure compensator valve <NUM>. This way, in an over-demand case where more than the flow capacity of the source <NUM> is requested, the amount of fluid flow provided to the actuators of the worksections <NUM>, <NUM> are reduced proportionally as opposed to providing more fluid flow to the worksection <NUM> with the lower load. As such, the configuration of the pressure compensator valves <NUM>-<NUM> allow for "flow-sharing" of the fluid provided by the source <NUM>. Further, in the over-demand case and for a given highest LS pressure level (P3) (the signal <NUM>), the inlet pressure level (P4) decreases because the total flow resistance of all the worksections increases. Full pump flow output cannot create the margin pressure (P4-P3). So, the pressure compensator valves move in a closing direction, thus reducing each workport output flow until each compensator reaches force equilibrium.

<FIG> illustrates a cross-sectional view of the worksection <NUM> that includes the pressure compensator valve <NUM>, in accordance with an example implementation. The main control valve <NUM> is a directional control valve having a main control spool <NUM> slidably accommodated (i.e., axially movable) in a longitudinal bore <NUM> in the worksection body <NUM>. <FIG> illustrates the main control spool <NUM> in a neutral position.

If the pilot valve <NUM> is actuated (e.g., via an electric command from an electronic controller), a pilot pressure signal is provided via pilot passage <NUM> to a first cavity <NUM> at a first end of the main control spool <NUM>. As a result, the main control spool <NUM> can shift in a first axial direction (e.g., distal direction to the right in <FIG>). If the pilot valve <NUM> is actuated (e.g., via an electric command from an electronic controller), a pilot pressure signal is provided via pilot passage <NUM> to a second cavity <NUM> at a second end of the main control spool <NUM>. As a result, the main control spool <NUM> can shift in a second axial direction (e.g., proximal direction to the left in <FIG>).

Inlet fluid flow from the source <NUM> of fluid is provided through the inlet flow passage <NUM> to a compensator inlet cavity <NUM> formed in the worksection body <NUM>. The fluid is then regulated by the pressure compensator valve <NUM> by flowing through a flow area <NUM> then to a regulated flow passage <NUM>. The size of the flow area <NUM>, and the pressure drop thereacross, changes based on axial position of a compensator spool <NUM> as described in detail below.

The main control spool <NUM> varies in diameter along its length to form lands of variable diameters capable of selectively interconnecting the various passages intercepting the longitudinal bore <NUM> to control flow of fluid to and from the workports <NUM>, <NUM>. The lands of the main control spool <NUM> cooperate with internal surfaces of the worksection body <NUM> to define variable metering orifices that allows fluid flow therethrough. For example, the main control spool <NUM> has land <NUM>, land <NUM>, land <NUM>, land <NUM>, land <NUM>, and land <NUM> configured to cooperate with the internal surfaces of the worksection body <NUM> to form the variable metering orifices and control the fluid flow rate and fluid direction through the worksection <NUM>. The variable metering orifices are spool-to-bore cylindrical area openings between the main control spool <NUM> and the internal surfaces of the worksection body <NUM> that form when the main control spool <NUM> shifts axially therein.

For instance, if the pilot valve <NUM> is actuated and the main control spool <NUM> shifts axially in the distal direction, the land <NUM> moves in the distal direction. As the land <NUM> moves distally to the extent that it moves past an edge of the internal surface of the worksection body <NUM> interfacing therewith, a metering orifice is formed that allows fluid flow from the regulated flow passage <NUM> to a metered flow passage <NUM> configured as a looped passage as depicted in <FIG>.

At the same time that the land <NUM> moves past the edge of the internal surface of the worksection body <NUM>, the land <NUM> can also move past a respective edge in the internal surface of the worksection body <NUM> and another metering orifice is formed that allows fluid flow from the metered flow passage <NUM> to a first workport passage <NUM>. The workport passage <NUM> is fluidly coupled to the workport <NUM> and thus fluid flows from the workport passage <NUM> to the workport <NUM>, and then to the actuator controlled by the worksection <NUM>. Fluid returning from the actuator through the workport <NUM> flows through a second workport passage <NUM>, then through another metering orifice formed between the land <NUM> and the internal surface of the worksection body <NUM> to a reservoir cavity <NUM>, which is fluidly coupled to the return flow passage <NUM> and the reservoir <NUM> depicted in <FIG>.

Conversely, if the pilot valve <NUM> is actuated and the main control spool <NUM> shifts axially in the proximal direction, the land <NUM> moves in the proximal direction. As the land <NUM> moves proximally to the extent that it moves past an edge of the internal surface of the worksection section <NUM>, a metering orifice is formed that allows fluid flow from the regulated flow passage <NUM> to the metered flow passage <NUM>.

At the same time that the land <NUM> moves past the edge of the internal surface of the worksection body <NUM>, the land <NUM> can also move past a respective edge in the internal surface of the worksection section <NUM> and another metering orifice is formed that allows fluid flow from the metered flow passage <NUM> to the workport passage <NUM>. The workport passage <NUM> is fluidly coupled to the workport <NUM> and thus fluid flows from the workport passage <NUM> to the workport <NUM>, and then to the actuator controlled by the worksection <NUM>. Fluid returning from the actuator through the workport <NUM> flows through the workport passage <NUM>, then through another metering orifice formed between the land <NUM> and the internal surface of the worksection body <NUM> to another reservoir cavity <NUM>, which is fluidly coupled to the return flow passage <NUM> and the reservoir <NUM> depicted in <FIG>.

As mentioned above, the pressure compensator valve <NUM> is configured to be subjected to fluid forces resulting from four pressure signals and the biasing force of the compensator spring <NUM>. The force balance between these forces determines the axial position of the compensator spool <NUM> within the worksection body <NUM> and thus determines the size of the flow area <NUM>. The pressure compensator valve <NUM> is configured such that the size of the flow area <NUM> varies based on variations in the forces to change the drop in pressure level across the flow area <NUM>. This way, the pressure drop in the pressure level of fluid as it flows from the regulated flow passage <NUM> to the metered flow passage <NUM> remains substantially constant.

<FIG> illustrates a partial cross-sectional view of the worksection <NUM> depicting the pressure compensator valve <NUM>, in accordance with an example implementation. <FIG> represents a zoom-in view of the cross-sectional view in <FIG> to illustrate construction details of the pressure compensator valve <NUM>.

The pressure compensator valve <NUM> comprises the compensator spool <NUM>, a first piston <NUM> disposed at a distal end of the compensator spool <NUM>, and a second piston <NUM> disposed in a cavity <NUM> formed at a proximal end of the compensator spool <NUM>. The first piston <NUM> can be referred to as a valve load-sense signal piston <NUM> as it is subjected to the signal <NUM> (valve LS pressure signal). The second piston <NUM> can be referred to as a pump signal piston or fluid source signal piston as it is subjected to the signal <NUM> from the source <NUM>. The compensator spool <NUM> is slidable about an exterior surface of the second piston <NUM>.

The first piston <NUM> is coupled to, and axially movable with, the compensator spool <NUM>. For example, the first piston <NUM> can have an exterior annular groove formed on an exterior peripheral surface thereof that is configured to engage with a slot or an interior annular groove formed in an interior surface of the compensator spool <NUM>. This configuration represents a T-slot engagement configuration; however, other arrangements can be implemented to mechanically couple or link the first piston <NUM> to the compensator spool <NUM>.

A proximal end of the second piston <NUM> is secured against a compensator plug <NUM> that is fixedly disposed in the worksection <NUM> adjacent the proximal end of the compensator spool <NUM>. The compensator spring <NUM> is disposed about the exterior surface of the compensator plug <NUM>. A proximal end of the compensator spring <NUM> is secured against a shoulder <NUM> formed in the exterior surface of the compensator plug <NUM> and a distal end of the compensator spring <NUM> rests against a proximal end of the compensator spool <NUM>. With this configuration, the compensators spring <NUM> applies a biasing force Fspr on the compensator spool <NUM> in the distal direction.

Referring to Figured <NUM>, <NUM>, and <NUM> together, regulated fluid flow in the regulated flow passage <NUM> provides the first signal <NUM> representing pressure level (P1) of regulated fluid flow provided from the pressure compensator valve <NUM> to the main control valve <NUM> of the worksection <NUM>. The signal <NUM> applies a first fluid force F1 on the compensator spool <NUM> in the proximal direction.

When the main control spool <NUM> shifts, the metered flow passage <NUM> is fluidly coupled to the workport <NUM> (via the workport passage <NUM>) or the workport <NUM> (via the workport passage <NUM>) based on the direction in which the main control spool <NUM> shifts. Thus, fluid in the metered flow passage <NUM> provides the second signal <NUM> representing local LS pressure signal having the pressure level (P2) of the workport <NUM> or the workport <NUM>. Fluid in the metered flow passage <NUM> is communicated through a channel <NUM> formed in the compensator spool <NUM> to a load-sense cavity <NUM>, and therefore applies a second fluid force F2 on an annular area of the proximal end of the compensator spool <NUM> in the distal direction.

The third signal <NUM> representing the highest LS pressure level (P3) transmitted via the valve LS passage <NUM> is communicated in the valve assembly <NUM> to valve load-sense cavity <NUM> formed in the worksection <NUM> at a distal end of the first piston <NUM>. Fluid in the valve LS cavity <NUM> applies a third fluid force F3 on the first piston <NUM>, and thus on the compensator spool <NUM> coupled thereto, in the proximal direction.

The fourth signal <NUM> representing inlet pressure level of fluid supplied from the source <NUM> through the inlet flow passage <NUM> is communicated to the compensator inlet cavity <NUM>. Fluid is then communicated from the compensator inlet cavity <NUM> to the cavity <NUM> through cross-hole <NUM>, channel <NUM>, and orifice <NUM> formed in the compensator spool <NUM>. Because the second piston <NUM> is secured against the compensator plug <NUM>, fluid in the cavity <NUM> applies a fourth fluid force F4 on the compensator spool <NUM> in the distal direction.

Notably, the second piston <NUM> has a spherical proximal surface <NUM> interfacing with the compensator plug <NUM>. Having such a spherical surface at the proximal end of the of the second piston <NUM>, as opposed to a flat surface, can reduce the likelihood of binding between the second piston <NUM> and the compensator plug <NUM> under pressure. Further, the spherical proximal surface <NUM> can also compensate for manufacturing tolerance issues such as out-of-flatness or eccentricities in the mating parts. Also, under fluid force from pressurized fluid, the second piston <NUM> can rotate along its longitudinal axis. The curvedness of the spherical proximal surface <NUM> can reduce flow hysteresis and non-repeatability of performance of the pressure compensator valve <NUM> that might result from such rotation.

A force balance between the biasing force Fspr of the compensator spring <NUM> and the four fluid forces F1-F4 determines an axial position of the compensator spool <NUM> and the size of the flow area <NUM>. A force equilibrium equation of forces acting on the compensator spool <NUM> can be expressed as follows:.

Thus, the difference between the forces F1 and F2, which represents the difference in pressure level (P1-P2) between pressure level in the regulated flow passage <NUM> and pressure level in the metered flow passage <NUM>, can be expressed as: <MAT>.

As a particular example for illustration, if the outer diameter of the pistons <NUM>, <NUM> is <NUM>,<NUM> (<NUM>,<NUM> inches) and the outer diameter of the compensator spool <NUM> is <NUM>,<NUM> (<NUM>,<NUM> inches), then the pressure difference (P1-P2) can be determined as: <MAT> Without the compensator spring <NUM>, the pressure difference (P1-P2) would be <NUM>( P<NUM> - P<NUM>).

Notably, because the difference in pressure level (P1-P2) is based on pressure the difference in pressure level (P4-P3), the axial position of the compensator spool <NUM> responds to, not only local LS pressure signal (i.e., the signal <NUM>, P2), but also the "global" LS signal (i.e., i.e., the signal <NUM>, P3). With this configuration, if the actuator controlled by the worksection <NUM> is subjected to a lower load compared to the actuator associated with another actuated worksection, i.e., if P2 < P3, the higher global LS signal (P3) can cause the first piston <NUM> to push the compensator spool <NUM> in the proximal direction. As a result, the flow area <NUM> is further restricted to reduce flow rate therethrough. As such, priority is not given to the worksection having the lowest LS pressure signal, but rather fluid flow rates are reduced proportionally for all actuated worksections.

Also, the bypass compensator valve <NUM> described above operates to maintain the pressure differential (P4-P3) substantially constant and equal to a particular pressure margin setting. Alternatively, if a load-sensing pump is used, the pump can operate to maintain the pressure margin setting, i.e., maintain the pressure differential (P4-P3) substantially constant. As a result of equation (<NUM>), the pressure differential (P1-P2), which the pressure difference between fluid in the regulated flow passage <NUM> upstream of the main control spool <NUM> and fluid in the metered flow passage <NUM> downstream of the main control spool <NUM>, can be maintained substantially constant.

Maintaining the pressure drop from the regulated flow passage <NUM> to the metered flow passage <NUM> across the variable metering orifice (formed when the main control spool <NUM> shifts) substantially constant facilitates proportionality between the command to the pilot valves <NUM>, <NUM> and the fluid flow rate to the workports <NUM>, <NUM>. Particularly, fluid flow through the variable metering orifice can be determined as <MAT>. K is a variable that is proportional to the size of the variable metering orifice formed when the main control spool <NUM> shifts, and thus K is determined based on the magnitude of the command signal to the pilot valve <NUM>, <NUM>. Thus, maintaining (P1-P2) substantially constant causes the command signal to the pilot valves <NUM>, <NUM> to be proportional to the flow rate Q. Further, as can be inferred from equation (<NUM>), the presence of the compensator spring <NUM> can increase the pressure difference (P1-P2), and can thus increase the fluid flow rate from the regulated flow passage <NUM> to the metered flow passage <NUM> for the worksection <NUM>.

The presence of the compensator spring <NUM> can provide several enhancements for the worksection <NUM>. As a first example enhancement, the output fluid flow from the worksection <NUM> to the workport <NUM> or the workport <NUM> and to the actuator fluidly coupled thereto can be varied by adjusting the biasing force of the compensator spring <NUM>.

Adjusting the biasing force of the compensator spring <NUM> can be implemented in various ways. As a first implementation, as shown in <FIG>, one or more shims <NUM> can be placed at a proximal end of the compensator spring <NUM>, and particularly between the proximal end of the compensator spring <NUM> and the shoulder <NUM>. Adding more shims reduces the initial length of the compensator spring <NUM>, thereby increasing the biasing force of the compensator spring <NUM> applied to the compensator spool <NUM>. Based on equation (<NUM>) above, increasing the biasing force of the compensator spring <NUM> can increase the pressure differential (P1-P2), thereby increasing the fluid flow rate through the worksection <NUM>.

Conversely, reducing the number of the shims <NUM> (or removing the shims <NUM>) relaxes the compensator spring <NUM> and increases its initial length, thereby decreasing the biasing force of the compensator spring <NUM> applied to the compensator spool <NUM>. Based on equation (<NUM>) above, decreasing the biasing force of the compensator spring <NUM> can decrease the pressure differential (P1-P2), thereby decreasing the fluid flow rate through the worksection <NUM>.

In a second implementation, adjusting the initial length and biasing force of the compensator spring <NUM> can be accomplished by using an adjustment piston. <FIG> illustrates a cross-sectional view of the worksection <NUM> with the pressure compensator valve <NUM> having an adjustment piston <NUM>, in accordance with an example implementation.

A distal end of the adjustment piston <NUM> interfaces with or contacts the proximal end of the compensator spring <NUM> such that longitudinal or axial motion of the adjustment piston <NUM> changes the length of the compensator spring <NUM>. The adjustment piston <NUM> can be threadedly coupled to a compensator plug <NUM> at threaded region <NUM>. The adjustment piston <NUM> can also be threadedly coupled to a nut <NUM> and can further be threadedly coupled to threads in the worksection body <NUM>.

If the adjustment piston <NUM> is rotated in a first rotational direction (e.g., clockwise), the adjustment piston <NUM> moves axially in the distal direction (e.g., to the right in <FIG>) due to its threaded engagement with the compensator plug <NUM> and the worksection body <NUM>. Movement of the adjustment piston <NUM> in the distal direction compresses the compensator spring <NUM> and increases its biasing force. Conversely, if the adjustment piston <NUM> is rotated in a second rotational direction (e.g., counter-clockwise), the adjustment piston <NUM> moves axially in the proximal direction (e.g., to the left in <FIG>). Movement of the adjustment piston <NUM> in the proximal direction relaxes the compensator spring <NUM> and decreases its biasing force.

In an example experiment, the main control spool <NUM> of the worksection <NUM> is shifted and the respective main control spool of the worksection <NUM> is also shifted to allow a fluid flow rate of about <NUM> GPM through the worksection <NUM>. Without shims being used, the fluid flow rate through the worksection <NUM> can be about <NUM>,<NUM> lpm (<NUM>,<NUM> gpm). When the shims <NUM> are added with a total thickness of about <NUM>,<NUM> (<NUM>,<NUM> inches) and a compensator spring with a higher spring rate is used to increase the spring force, the flow rate through the worksection <NUM> increased to <NUM> lpm (<NUM>,<NUM> gpm). When the worksection <NUM> is actuated without actuating the worksection <NUM>, a flow rate increase from <NUM>,<NUM> lpm (<NUM>,<NUM> gpm) (without shims) to <NUM>,<NUM> lpm (<NUM> gpm) is observed when the shims <NUM> having a thickness of <NUM>,<NUM> (<NUM>,<NUM> inches) are added and the compensator spring with the higher spring rate is used.

As a second example enhancement, the compensator spring <NUM> can enhance compensation efficiency of the pressure compensator valve <NUM>. Efficiency of a pressure compensator valve is related to its ability to maintain pressure compensation (i.e., maintain the pressure differential P1-P2 substantially constant) regardless of the changes in the different pressure levels P1, P2, P3, and P4, changes in flow rates across the flow area <NUM>, and the resulting changes in flow forces acting on the compensator spool <NUM>.

Bernoulli flow forces can result from accelerating fluid mass through the flow area <NUM> between the compensator spool <NUM> and the internal surfaces of the worksection body <NUM>. The flow forces can have an axial component that acts on the compensator spool <NUM> in a closing (proximal) direction (e.g., to the left in <FIG>) opposing the opening force applied to the compensator spool <NUM> (e.g., the opening force resulting from the signals <NUM> (P2) and the signal <NUM> (P4) that act to move the compensator spool <NUM> in the distal direction to increase the size of the flow area <NUM>). In some cases, the flow forces are sufficiently high that the opening forces might not be sufficient to maintain pressure compensation (i.e., maintain the pressure differential P1-P2 substantially constant) as the pressure level P4 increases or the flow rate across the flow area <NUM> increases.

Presence of the compensator spring <NUM> can enhance compensation efficiency of the pressure compensator valve <NUM>, particularly in high flow rate ranges through the flow area <NUM>. In particular, the biasing force of the compensator spring <NUM> acts on the compensator spool <NUM> in the distal direction to counter the flow force acting on the compensator spool <NUM> in the proximal direction. This way, the size of the flow area <NUM> may remain sufficient to maintain the flow rate thereacross independent of inlet or workport load pressure changes (i.e., changes in any of P1, P2, P3, P4).

<FIG> illustrates a graph <NUM> associated with experimental results showing compensation efficiency when the compensator spring <NUM> is not used, in accordance with an example implementation. The left y-axis in the graph <NUM> represents pressure level P4 (the signal <NUM>) in the inlet flow passage <NUM>, whereas the right y-axis in the graph <NUM> represents the fluid flow rate Q provided to the workport <NUM> of the worksection <NUM> when the main control spool <NUM> is shifted in the proximal direction (e.g., to the left in <FIG>). The flow rate Q provided to the workport <NUM> is the same as the flow rate provided from the regulated flow passage <NUM> to the metered flow passage <NUM>.

Line <NUM> represents variation in the pressure level P4 as the highest LS pressure (the pressure level P3) changes. As the pressure level P4 increases, the pressure compensator valve <NUM> operates to maintain the pressure level P1 to maintain a particular pressure differential (P1-P2) (e.g., <NUM>,<NUM> bar (<NUM> psi)). Line <NUM> illustrates variation in the flow rate Q through the workport <NUM> as the pressure level P4 changes.

The flow rate Q should remain substantially constant for a given axial main control spool position as the pressure level P4 changes because the pressure differential (P1-P2) remains constant. However, the line <NUM> illustrates that the flow rate Q decreases from <NUM>,<NUM> lpm to <NUM>,<NUM> lpm (<NUM> gpm to <NUM> gpm) (a <NUM>% reduction in flow rate) as P4 increases from about <NUM>,<NUM> bar (<NUM> psi) to about <NUM>,<NUM> bar (<NUM> psi). Such reduction likely results from the flow forces that increase and tend of reduce the size of the flow area <NUM> as the pressure level P4 increases.

<FIG> illustrates a graph <NUM> associated with experimental results showing enhancement in compensation efficiency when the compensator spring <NUM> is used, in accordance with an example implementation. In this experiment, the compensator spring <NUM> used has a spring rate of about <NUM> pound per inch resulting in a spring force of about <NUM> pound-force. Line <NUM> illustrates variation in the flow rate Q through the workport <NUM> as the pressure level P4 changes when the compensation spring <NUM> is used.

As shown in <FIG>, when the compensation spring <NUM> is used, the line <NUM> shows the flow rate Q decreases from <NUM>,<NUM> lpm (<NUM> gpm) to about <NUM>,<NUM> lpm (<NUM>,<NUM> gpm) (a <NUM>,<NUM> lpm (<NUM>,<NUM> gpm) reduction in flow rate as opposed to <NUM>,<NUM> lpm (<NUM> gpm) when the compensator spring <NUM> is not used). As such, the compensator spring <NUM> enhances the compensation efficiency of the pressure compensator valve <NUM> by about <NUM>%.

A third example enhancement involves using the compensator spring <NUM> to adjust flow split between the worksections <NUM>-<NUM> when the total amount of flow rate requested by all worksections exceeds the flow capacity of the source <NUM> (e.g., an over-demand case where total flow rate requested exceeds a flow capacity of a pump). The compensator spring <NUM> can be used to provide a flow priority to a particular worksection (e.g., the worksection <NUM>) or can be used to equalize flow rates among the worksections.

<FIG> illustrates a graph <NUM> associated with experimental results showing flow rate split between the worksections <NUM>, <NUM> when the compensator spring <NUM> is not used, in accordance with an example implementation. The x-axis represents percentage of maximum electric command to the pilot valve <NUM> of the worksection <NUM> and the y-axis represents fluid flow rate through the worksections <NUM>, <NUM> in GPM. Line <NUM> represents variation in flow rate Q1 through the worksection <NUM> and line <NUM> represents variation in flow rate Q2 through the worksection <NUM> as the magnitude of the command signal to the pilot valve <NUM> changes.

In this experiment the maximum available flow rate is set to <NUM>,<NUM> lpm (<NUM> gpm). Initially, the main control spool of the worksection <NUM> is actuated with a maximum command (e.g., <NUM>% pulse width modulated command to the respective pilot valve) while the main control spool <NUM> of the worksection <NUM> is unactuated. In this state, as indicated with the line <NUM>, the maximum available flow rate of <NUM>,<NUM> lpm (<NUM> gpm) is provided through the worksection <NUM>.

When the command signal to the pilot valve <NUM> of the worksection <NUM> reaches about <NUM>% of the maximum command and increases therefrom to about <NUM>% of maximum command, the flow rate Q1 through the worksection <NUM> increases. The command to the pilot valve of the worksection <NUM> remains at maximum command, and therefore this state simulates a flow over-demand case where the requested flow exceeds the available <NUM>,<NUM> lpm (<NUM> gpm). In response to such over-demand state, as the flow rate Q1 increases from <NUM> to about <NUM>,<NUM> lpm (<NUM> gpm), the flow rate Q2 through the worksection <NUM> decreases from <NUM>,<NUM> lpm to about <NUM>,<NUM> lpm (<NUM> gpm to about <NUM> gpm). However, the worksection <NUM> receives a higher flow rate compared to the worksection <NUM>.

In some examples, it may be desirable to assign the worksection <NUM> flow priority over the worksection <NUM>. The compensator spring <NUM> can be added to the pressure compensator valve <NUM> of the worksection <NUM> to achieve such flow priority.

<FIG> illustrates a graph <NUM> associated with experimental results showing flow rate split between the worksections <NUM>, <NUM> when the compensator spring <NUM> is used, in accordance with an example implementation. Line <NUM> illustrates variation in the flow rate Q1 through the worksection <NUM> and line <NUM> represents variation in flow rate Q2 through the worksection <NUM> as the magnitude of the command signal to the pilot valve <NUM> changes. The conditions of the test remains the same as the test associated with the graph <NUM> described above except that the worksection <NUM> now has the compensator spring <NUM>.

As shown by the lines <NUM>, <NUM>, when the command signal to the pilot valve <NUM> exceeds about <NUM>% of maximum command, the flow rate Q1 increases while the flow rate Q2 decreases. In contrast to the graph <NUM>, however, the flow rate Q1 through the worksection <NUM> reaches <NUM>,<NUM> lpm (<NUM>,<NUM> gpm) whereas the flow rate Q2 of the worksection <NUM> decreases to about <NUM>,<NUM> lpm (<NUM>,<NUM> gpm). a As such, the worksection <NUM> has flow priority over the worksection <NUM> due to using the compensator spring <NUM> in the pressure compensator valve <NUM> of the worksection <NUM>.

<FIG> is a flowchart of a method <NUM> for operating the pressure compensator valve <NUM>, in accordance with an example implementation.

The method <NUM> shown in <FIG> presents an example of a method that can be used with the valve assembly <NUM>, the hydraulic system <NUM>, and the pressure compensator valve <NUM> shown throughout the Figures, for example. The method <NUM> may include one or more operations, functions, or actions as illustrated by one or more of blocks <NUM>-<NUM>. Although the blocks are illustrated in a sequential order, these blocks may also be performed in parallel, and/or in a different order than those described herein. Also, the various blocks may be combined into fewer blocks, divided into additional blocks, and/or removed based upon the desired implementation. It should be understood that for this and other processes and methods disclosed herein, flowcharts show functionality and operation of one possible implementation of present examples. Alternative implementations are included within the scope of the examples of the present disclosure in which functions may be executed out of order from that shown or discussed, including substantially concurrent or in reverse order, depending on the functionality involved, as would be understood by those reasonably skilled in the art.

At block <NUM>, the method <NUM> includes applying a first fluid force on the compensator spool <NUM> of the pressure compensator valve <NUM>, wherein the pressure compensator valve <NUM> is disposed in the worksection body <NUM> of the worksection <NUM>, the worksection body <NUM> having (i) the inlet cavity <NUM> configured to receive fluid from the source <NUM> of fluid, (ii) the regulated flow passage <NUM>, and (iii) the metered flow passage <NUM>, and wherein the first fluid force is applied by fluid in the regulated flow passage <NUM>.

At block <NUM>, the method <NUM> includes applying a second fluid force on the compensator spool <NUM> by fluid in the metered flow passage <NUM>.

At block <NUM>, the method <NUM> includes applying a third fluid force on the compensator spool <NUM> by a load-sense signal received in the valve load-sense cavity <NUM> within the worksection body <NUM>.

At block <NUM>, the method <NUM> includes applying a fourth fluid force on the compensator spool <NUM> by fluid from the inlet cavity <NUM>.

At block <NUM>, the method <NUM> includes applying a biasing force on the compensator spool <NUM> by the compensator spring <NUM>, wherein an axial position of the compensator spool <NUM> is based on a force equilibrium between the first fluid force, the second fluid force, the third fluid force, the fourth fluid force, and the biasing force of the compensator spring <NUM>.

Claim 1:
A pressure compensator valve (<NUM>) comprising:
a compensator spool (<NUM>);
a first piston (<NUM>) coupled to a distal end of, and axially-movable with, the compensator spool (<NUM>);
a second piston (<NUM>) disposed in a cavity (<NUM>) formed at a proximal end of the compensator spool (<NUM>), wherein the compensator spool (<NUM>) is axially movable relative to the second piston (<NUM>);
a compensator spring (<NUM>) applying a biasing force on the compensator spool (<NUM>)
a regulated flow passage (<NUM>) configured to have fluid applying a first fluid force on the compensator spool (<NUM>) in a proximal direction;
a metered flow passage (<NUM>) configured to have fluid applying a second fluid force on the compensator spool (<NUM>) in a distal direction;
a load-sense cavity (<NUM>) configured to receive a load-sense signal applying a third fluid force on the first piston (<NUM>) and the compensator spool (<NUM>) coupled thereto in the proximal direction; and
a compensator inlet cavity (<NUM>) configured to receive inlet fluid from a source (<NUM>) of fluid, the inlet fluid applying a fourth fluid force on the compensator spool (<NUM>) in the distal direction, wherein an axial position of the compensator spool (<NUM>) is based on a force equilibrium between the first fluid force, the second fluid force, the third fluid force, the fourth fluid force, and the biasing force of the compensator spring (<NUM>).