Torque control system for a variable displacement pump

The present invention relates to a hydraulic pump system including a variable displacement pump that generates an outlet pressure. The hydraulic pump system also includes a control system that decreases a displacement volume of the variable displacement pump in response to an increase in the outlet pressure and increases a displacement volume of the variable displacement pump in response to a decrease in the outlet pressure.

TECHNICAL FIELD

The present disclosure relates generally to hydraulic systems. More particularly, the present disclosure relates to hydraulic systems including variable displacement pumps.

BACKGROUND

Hydraulic systems are used to transfer energy using hydraulic pressure and flow. A typical hydraulic system includes one or more hydraulic pumps for converting energy/power from a power source (e.g., an electric motor, a combustion engine, etc.) into hydraulic pressure and flow used to provide useful work at an actuator or other device (i.e., a load). A typical hydraulic pump includes a rotating group that includes one or more pistons carried within cylinders defined by a rotor coupled to an input shaft. The input shaft supplies torque for rotating the rotating group. As the rotating group rotates about a central axis of the input shaft, the pistons reciprocate (i.e., stroke) within the cylinders of the rotating group. This causes hydraulic fluid to be drawn into an input port of the pump and discharged from an output port of the pump. In a variable displacement pump, the volume of fluid displaced by the pump for each rotation of the rotating group (i.e., the displacement volume of the pump) can be varied to match hydraulic pressure and flow demands corresponding to the load. Typically, the displacement volume of a pump is varied by varying the stroke length of the pistons of the rotating group within their corresponding cylinders. The workload experienced by hydraulic pumps is dependent upon factors such as the working pressure and the pump output flow. In some operating conditions, the torque required to drive the pump to satisfy a given workload may exceed the capacity of the power source.

SUMMARY

One aspect of the present disclosure relates to a torque control system for a variable displacement pump that reduces the pump output flow when the driving effort reaches a threshold set by the torque control system thereby preventing the power source from overloading.

Another aspect of the present disclosure relates to a torque control system for a variable displacement pump that decreases a stroke length of the variable displacement pump in response to an increase in pump outlet pressure and increases the stroke length of the variable displacement pump in response to a decrease in pump outlet pressure.

A further aspect of the present disclosure relates to a hydraulic pump system including a variable displacement pump that generates an outlet pressure. The system also includes a control system that decreases a displacement volume of the variable displacement pump in response to an increase in the outlet pressure and increases a displacement volume of the variable displacement pump in response to a decrease in the outlet pressure.

Other aspects of the present disclosure relates to a control system for a variable displacement pump having a torque control function that automatically adjusts the pump displacement in response to load pressure. In certain examples, the control system reduces displacement at higher pressures to limit the input torque demand. In this way, a torque limit is maintained across a range of operating pressures, speeds and oil temperature. This use of torque control allows for higher flow at low pressure while maintaining the ability to achieve high pressure without exceeding the torque capacity of the power source (e.g., motor or engine) driving the pump.

Other aspects of the present disclosure relate to variable displacement pump controlled by control system including torque control valve in which a spring preload alone of the torque control valve governs a torque limit of the pump.

DETAILED DESCRIPTION

FIG. 1illustrates a variable displacement pump system20in accordance with the principles of the present disclosure. The variable displacement pump system20includes a variable displacement pump22controlled by a pump control system23. The pump control system23includes a valve stack25having a pressure compensation valve arrangement24and a torque control valve26. The pump control system20also includes a control piston28for controlling a position of a swash plate48of the variable displacement pump22.FIGS. 2-4are various cross-sectional views showing how the control piston28interfaces with the swash plate48.FIG. 5is a cross-sectional view taken through the valve stack25.

As best shown atFIG. 4, the variable displacement pump22includes a control piston sleeve32that is mounted within a control piston cylinder35defined by a housing30of the pump22. The control piston sleeve32defines a bore33in which the control piston28is mounted. Although it is primarily described in the present disclosure that the sleeve32is mounted within the control piston cylinder35defined by the housing30of the pump22, it is also possible that the sleeve32is formed to be integral with the pump housing30. In yet other embodiments, the housing30of the variable displacement pump22is configured to define the bore33without the control piston sleeve32. For example, at least a portion of the control piston cylinder35is configured to replace the control piston sleeve32so that the control piston28is mounted directly within the control piston cylinder35(i.e., the control piston cylinder35defines the bore33without the control piston sleeve32). In this configuration, the control piston cylinder35, and/or at least a portion of the pump housing30that is associated with the control piston cylinder35, includes features corresponding to the features of the control piston sleeve32as described in the present disclosure.

With continued reference toFIG. 4, the variable displacement pump22includes a rotating group34mounted within the pump housing30. The rotating group34includes a rotor36defining a plurality of piston cylinders38that receive pistons40. The variable displacement pump22also includes an input shaft42that defines an axis of rotation44. The input shaft42is coupled to the rotor36such that torque can be transferred from the input shaft42to the rotor36thereby allowing the input shaft42and the rotor36to rotate together about the axis of rotation44. In certain examples, a splined connection can be provided between the input shaft42and the rotor36. As depicted, bearings46are provided between the input shaft42and the pump housing30for allowing the input shaft42to rotate relative to the pump housing30about the axis of rotation44.

Still referring toFIG. 4, the swash plate48is also positioned within the pump housing30. The swash plate48is pivotally movable relative to the axis of rotation44between a neutral position (seeFIGS. 3, 4, and 6) and a maximum displacement position (seeFIGS. 2 and 7). The neutral position can also be referred to as a minimum displacement position. It will be appreciated that movement of the swash plate48varies an angle of swash plate48relative to the axis of rotation44. Varying the angle of the swash plate48relative to the axis of rotation44varies the displacement volume of the variable displacement pump22. The displacement volume is the amount of hydraulic fluid displaced by the variable displacement pump22for each rotation of the rotating group34. When the swash plate48is in the neutral position, the pump displacement has a minimum value. In certain examples, the minimum value can be zero displacement. When the swash plate48is in the maximum displacement position, the variable displacement pump22has a maximum displacement value.

Referring still toFIG. 4, the pistons40of the rotating group34include cylindrical heads50on which hydraulic shoes52are mounted. The hydraulic shoes52have end surfaces54that oppose the swash plate48. Typically, hydraulic fluid provides a hydraulic bearing layer between the end surfaces54and the swash plate48that facilitates rotating the rotating group34about the axis of rotation44relative to the swash plate48. When the swash plate48is in the neutral position, the swash plate is generally perpendicular relative to the axis of rotation44thereby causing a stroke length of the pistons40within their respective piston cylinders38to be at or near zero. By adjusting the angle of the swash plate48relative to the axis of rotation44, the stroke length of the pistons40within their corresponding piston cylinders38is adjusted. When the swash plate48is positioned at a non-perpendicular angle relative to the axis of rotation44, the pistons reciprocate one stroke length for each rotation of the rotor36about the axis of rotation44. The stroke length increases as the swash plate48is moved from the neutral position toward the maximum displacement position. As the pistons40reciprocate within their corresponding piston cylinders38, the rotating group34provides a pumping action that draws hydraulic fluid into an inlet56(see schematically atFIG. 8) of the variable displacement pump22and forces hydraulic fluid out of an outlet58(see schematically atFIG. 8) of the variable displacement pump22.

The control piston28is used to control the position or angle of the swash plate48relative to the axis of rotation44. The control piston28includes a first end60and an opposite second end62. The first end60of the control piston28is shown engaging the swash plate48. A spring64is provided within the pump housing30for biasing the swash plate48toward the maximum displacement position. The angle of the swash plate48relative to the axis of rotation44is adjusted by moving the control piston28axially within the sleeve32(or the control piston cylinder35where the system20is configured without the sleeve32). In certain examples, a control pressure is applied to the second end62of the control piston28to cause the control piston28to move the swash plate48from maximum displacement position toward the neutral position. The force applied by the control pressure to the second end62of the control piston28must exceed the spring force of the spring64and other forces to move the swash plate48from the maximum displacement position toward the neutral position. Such other forces include hydraulic forces introduced by the pressures within the piston cylinders38and transmitted to the swash plate48via the pistons40and through the shoes52. When the force applied to the second end62of the swash plate control piston28is less than a combination of the spring force of the spring64and the other forces, the combination of the forces moves the swash plate48back towards the maximum displacement position.

It will be appreciated that the control system of the variable displacement pump22can provide a torque control function. In certain examples, various elements can cooperate to provide the torque limiting function of the pump. In one example, the torque control valve26and the control piston28can cooperate to provide the torque limiting function. In certain examples, the torque control valve26can function similar to a load sense or pressure compensator valve, and the control piston28can include an integrated hydraulic potentiometer that generates a torque limiting pressure signal Ptcwhich interfaces with the torque control valve26to provide a pressure balancing function with respect to a spool66of the torque control valve26.

The control piston28having an integral potentiometer29is shown atFIG. 9. Referring toFIG. 9, the control piston28includes a first zone68positioned nearest the first end60of the control piston28and a second zone70positioned nearest the second end62of the control piston28. The control piston28also includes a third zone72positioned between the first and second zones68and70. In certain examples, the first and second zones70have generally smooth, cylindrical surfaces. In certain examples, the third zone72has an integrated structure that can function as the hydraulic potentiometer29. In certain examples, the structure can include a helical groove31(e.g., similar to a thread), that permits the passage of laminar flow across the third zone72between the first and second zones68,70. The hydraulic pressure of fluid passing through the third zone72along the helical groove will decrease in a linear manner from one end to the other of the helical groove. The hydraulic pressure along the second zone70can be generally the same throughout and similarly the hydraulic pressure along the first zone68can be generally the same throughout. In certain examples, the hydraulic pressure provided to the first zone68is case pressure of the pump housing (i.e., essentially tank/reservoir/drain pressure).

In certain examples, the second zone70can be in fluid communication with the outlet58(schematically shown inFIG. 8) of the pump22so as to be generally at outlet pressure. Thus, under many operating conditions, the hydraulic pressure at the second zone70is substantially higher than the hydraulic pressure at the first zone68. This causes hydraulic fluid to flow along the helical groove at the third zone72such that the hydraulic fluid flows along a helical path that extends circumferentially around the control piston28as the path extends axially along the length of the control piston28. When the hydraulic fluid flows along the helical groove, the pressure of the hydraulic fluid will decrease in a linear manner as the hydraulic fluid flows from the second zone70toward the first zone68.

Referring toFIG. 10, the torque control valve26of the variable displacement pump22includes a valve body80defining a bore82in which a valve spool66is mounted. The valve body80also defines a spring chamber84containing a spring86. The spring86applies a spring force in a first direction88to the spool66(i.e., a pre-load). In certain examples, the spring load or force corresponding to the spring86sets a maximum torque limit of the variable displacement pump22. In certain examples, the torque control valve26can include a spring pre-load adjustment mechanism90which allows the spring pre-load of the spring86to be manually adjusted. In certain examples, the spring pre-load adjustment mechanism90includes a threaded member91(i.e., a bolt or screw) that can be turned to adjust the spring pre-load and thereby adjust the torque limit of the pump. In certain examples, the spring pre-load adjustment mechanism90allows the torque setting of the pump to be adjusted without any disassembly.

In certain examples, the torque control valve26includes a first port94in fluid communication with the second end62of the control piston28, a second port96in fluid communication with the outlet58of the pump22and a third port98in fluid communication with the potentiometer29of the control piston28. The first port94provides control pressure to the second end62of the control piston28. Another port97is in fluid communication with tank pressure.FIG. 13shows an example fluid connection arrangement between the control piston28and the control valve26.

It will be appreciated that the spool66is configured to move axially within the bore82. Some embodiments of the spool66can be subdivided into two or more individual parts (e.g.,66A and66B inFIG. 10). In some embodiments, at least one of the individual parts of the spool66can be of different diameters. In other embodiments, the individual parts of the spool66can have the same diameter. Opposing axial forces are applied to opposite first and second ends92,93of the spool66to control the axial position of the spool66within the bore82. For example, the spring force from spring86as well as pressure within the spring chamber84cooperate to apply a first axial force F1to the first end92of the spool66. The pressure within the spring chamber84is determined by a signal pressure Ptcreceived from the potentiometer of the control piston28. The force applied to the spool66by the signal pressure Ptccan be referred to as a signal pressure force. A second force F2is applied to the second end93of the spool66. The second force F2is generated by the outlet pressure of the pump applied against the second end93of the spool66. This force F2can be referred to as an outlet pressure force. The forces F1and F2are opposite each other and can dynamically change toward a balanced condition.

Referring toFIG. 13, when the force F2exceeds the force F1(typically as a result of increased outlet pressure), the spool66is caused to move to the left thereby opening fluid communication between the port96and the port94. In this way, pump outlet pressure from the port96is provided to the port94. The increased pressure provided by the pump outlet pressure increases the control pressure provided to the second end62of the control piston28causing the control piston28to move the swash plate toward the neutral position thereby reducing the stroke length of the hydraulic pump22such that the pump displacement is decreased. Movement of the control piston28toward the neutral position combined with the increase in outlet pressure causes the magnitude of the signal pressure provided to the spring chamber84from the potentiometer29to increase thereby causing the force F1to increase to a point where F1exceeds F2and the spool moves back to the right thereby closing fluid communication between the ports96,94. The system operates such that the forces F1and F2iteratively adjust toward a re-balanced condition.

When the force F2falls below the force F1(typically as a result of decreased outlet pressure), the spool66moves to the right to a position where the port94is placed in fluid communication with tank pressure via port97thereby reducing the magnitude of the control pressure provided to the second end62of the control piston28. This reduction in control pressure causes the control piston28to allow the swash plate to be spring biased back toward the maximum displacement position such that the stroke length of the pistons is increased to increase the displacement volume of the pump. Movement of the control piston28toward the maximum displacement position combined with the decrease in outlet pressure causes the magnitude of the signal pressure Ptcprovided to the spring chamber84from the potentiometer29to decrease thereby causing the force F1to decrease to a point where F1is less than F2and the spool moves back to the left thereby closing fluid communication between the ports97,94. The system operates such that the forces F1and F2iteratively adjust toward a re-balanced condition.

Referring toFIGS. 11-13, the sleeve32(or the control piston cylinder35where the system20is configured without the sleeve32) that receives the control piston28defines an annulus100or other volume (i.e., a signal pressure output location) in fluid communication with the third port98of the control valve26. The annulus100is positioned at the interior of the sleeve32(or the control piston cylinder35) and opposes the exterior surface of the control piston28. In some embodiments, the annulus100on the exterior of the sleeve32(or the control piston cylinder35) is in fluid communication with the interior of the sleeve32(or the control piston cylinder35) through a plurality of passages103. When the control piston28is in the maximum displacement position, the annulus100and the passages103are positioned adjacent the interface between the first zone68and the third zone72. In contrast, when the control piston is in the neutral position, the annulus100and the passages103are positioned adjacent the interface between the third zone72and the second zone70. The magnitude signal pressure output from the signal pressure output location to the spring chamber84generally reduces linearly as the control piston28moves toward the maximum displacement position and increases linearly as the control piston moves toward the neutral position.

In certain examples, the sleeve32(or the control piston cylinder35where the system20is configured without the sleeve32) can also define an internal annulus102or volume/space at the second zone70that is in fluid communication with the pump outlet. In this way, the region of the sleeve32(or the control piston cylinder35) surrounding the second zone70can be provided at pump outlet pressure. In contrast, the region of the sleeve32(or the control piston cylinder35) surrounding the first zone68can be provided at case or tank pressure. In this way, when the control piston28is in the maximum displacement position, case or tank pressure is provided to the internal annulus100. Thus, the signal pressure output from the potentiometer29corresponds to case or tank pressure and is provided to the spring chamber84through the third port98. In contrast, when the control piston is in the neutral position, pump outlet pressure from the internal annulus102is provided to the internal annulus100. In this way, the signal pressure output from the potentiometer29corresponds to pump pressure and is provided to the spring chamber84through the third port98.

As the control piston28moves between the neutral position and the maximum displacement position, the hydraulic pressure provided to the internal annulus100varies linearly with the position of the control piston28since the pressure within the helical groove defined by the control piston28decreases in a linear manner from one end to the other. The pressure provided to the annulus100is thus dependent upon where the annulus100aligns with the third zone72. When the annulus100aligns with a first end of the third zone72, the hydraulic pressure provided to the annulus100is generally pump outlet pressure. When the annulus100aligns with the second end of the helical groove, the hydraulic pressure is generally case pressure (i.e., tank or drain pressure). In the region between the first and second ends of the helical groove, the hydraulic pressure provided to the annulus100varies linearly from outlet pressure to tank pressure.

In certain examples, the pump control system can be provided with minimum and maximum displacement limit features. In certain examples, the pump will only operate between the minimum and maximum displacements, regardless of operating conditions. This feature can override all other controls such as pressure compensator controls, load sense controls and torque controls.

In one example, the minimum displacement feature can be accomplished by adding a pressure relief passage120(seeFIG. 14) to the control piston28. The pressure relief passage120can have a side opening122(i.e., a blow hole) that is placed at a desired axial position along the axial length of the control piston28. When the side opening122is placed in fluid communication with case pressure, pressure applied to the second end of the control piston28is relieved to control pressure to case\tank thereby reducing the control pressure and preventing the pump from de-stroking further. For example, when the pressure relief hole122moves past the end of the sleeve32(or the control piston cylinder35where the system20is configured without the sleeve32), the second end of the piston28is placed in fluid communication with case pressure thereby reducing the control pressure and stopping movement of the piston that would cause further de-stroking of the pump. Thus, the pressure relief hole122in combination with the end of the sleeve32(or the control piston cylinder35) functions as a stop.

In certain examples, a maximum displacement feature can include an adjustable actuator such as an adjustment screw124that can determine the maximum displacement position of the control piston28within the sleeve32(or the control piston cylinder35where the system20is configured without the sleeve32). In certain examples, the maximum displacement adjustment mechanism can include a stop against which the control piston abuts when in the desired maximum displacement position. By adjusting the axial position of the stop within the sleeve32(or the control piston cylinder35), the maximum displacement of the pump can be adjusted.

It will be appreciated that the signal pressure provided from the hydraulic potentiometer of the control piston28varies with the position of the control piston28within the sleeve32(or the control piston cylinder35). For example, the value of the signal pressure Ptcprovided to the spring chamber84increases as the control piston28moves from the maximum displacement position toward the minimum displacement position. In this way, as the force F2increases with increased pump outlet pressure, the force F1also increases to counterbalance the force F2. As indicated previously, the force F1is the combined force applied to the spool66by the spring86and by the signal pressure within the spring chamber84. In this way, a force balanced relationship can be maintained with respect to the spool66in an axial orientation as the outlet pressure raises and lowers.

In certain examples, a torque control function is provided through the cooperation of two primary elements including a control valve and a hydraulic potentiometer. The control valve can be constructed like a load sense or pressure compensator valve, the features of which are known to those skilled in the art. The hydraulic potentiometer generates a torque limiting signal pressure, Ptcwhich is directed to the control valve spring chamber.

As indicated above, the control valve can be constructed like a standard load sense or pressure compensator valve except a signal pressure, Ptc, is applied to the spring chamber rather than case pressure or load sense pressure. When the force from pump outlet pressure, Pout, acting on the right area of the spool of the control valve is higher than the summation of signal pressure, Ptc, acting on the left area, A, of the spool plus spring preload force, Fs, the control valve ports pressure/flow to the control piston to de-stroke the pump.

When the torque control is active, it seeks to balance forces across the spool:
Pout=Ptc+Fs/A

The control valve is arranged in a hydraulically parallel circuit with the pressure compensation and load sense to allow override from pressure comp or load sense.

The hydraulic potentiometer generates a signal pressure, Ptc, that is the product of pump outlet pressure, Pout, and pump displacement, D. This pressure is ported to the spring chamber of the control valve. Ptcincreases proportional to Poutand decreases proportional to displacement, D:

Expressed in closed form:
Ptc=Pout(1−D)

When the hydraulic potentiometer provides a signal pressure according to the above relationship, a constant torque limit is achieved. This function is accomplished through two primary components: The control piston (moving with the swash plate) and the sleeve (connected to the housing) in which the piston translates.

With regard to the hydraulic potentiometer, the spiral groove feature creates as a long, narrow, ‘pipe’—connecting Poutto Ptank(zero gage pressure). Along this pipe, the pressure drop from Poutto Ptctankis linear along the length of the pipe. A spiral groove is preferable to create the ‘pipe’ feature, rather than a fixed clearance annular leak or straight axial groove(s), as it is much more robust in providing a linear signal (critical to torque limit accuracy) when dealing with manufacturing tolerances. The spiral feature is robust to variation in annular clearance between the piston and the bore as well as eccentricity and tilting of the piston within the bore.

With regard to the control piston28, the annular clearance between the housing bore (i.e., the sleeve bore) and the piston OD is very small compared to the cross-sectional area of the spiral, so the vast majority of the flow is along the spiral path. Since the spiral wraps around the piston many times the ‘pipe’ length is quite long, which creates a low flow situation and allows for a consistent pressure drop when manufacturing tolerances are considered.

The signal pressure, Ptc, picks up the pressure at a point along the piston that is fixed relative to the housing/sleeve. Its position and the axial length and start/end positions of the spiral feature are arranged such that it reads outlet pressure at zero displacement and tank pressure at full displacement. As the control piston moves linearly with pump displacement, the signal pressure reads pressure linearly to displacement (for a given Ptc).

Since the signal pressure reads pressure along the ‘pipe’ linearly with displacement (1−D), and the pressure along the ‘pipe’ is scaled linearly to outlet pressure, Pout, and length along the pipe, the resultant signal pressure provides the desired form: Ptc=Pout(1−D).

The design intent of a torque control is to limit torque to a constant value, independent of the state of pressure, displacement, pump speed, oil temperature, and torque setting. The basic equation for pump torque, T:
T=PoutD

Review of the 2 primary elements of the control:

1. The control valve acts to maintain a relationship pressure balance such that:

2. The hydraulic potentiometer creates a signal pressure:
Ptc=Pout(1−D)
Substituting Ptcfrom equations 2 into equation 1:

Pout=Pout⁡(1-D)+FsA
Expand and reduce:

Pout=Pout-Pout⁢D+FsAPout⁢D=FsA
And since

An equation for constant torque is thus derived. Since A is fixed, the torque limit is governed by the spring preload alone, Fs, meeting the design intent.

In reality, T=PoutD+Tlosswhere Tlossare inherent mechanical losses in the pump, which have some dependency on pressure, displacement, speed, temperature. Also, in reality Ptc=(Pout−Pml+Ptank)(1−D) where:a. Case pressure, Ptank, is slightly higher than pump inlet pressure.b. PmlAre “minor losses” that occur due to the flow transitioning from moving slower over a large area to moving faster in a small area, as well as turning corners (transition from the piston annulus to the spiral groove).

These additional losses would create a significant deviation from a constant torque limit (poor accuracy) if they were ignored. Therefore, adjustments can be made to the ‘ideal’ design of the valve, sleeve, and piston. These adjustments can mitigate the loss effects—providing a nearly constant torque limit that is robust to manufacturing tolerances and operating conditions. These parameters can be optimized through simulation and test iterations.

As shown previously, the torque control is governed by the equation:
T=Fs/A

Since the setting is determined by spring preload (as the control valve is active), the only adjustment needed is to adjust the control valve screw: Turn in to increase torque limit; turn out decrease torque limit. A single control setup covers all setting variations (for example, a range from about 20% to about 90% of max.).

A feature can exist in the pump to provide a minimum displacement limit. In this situation, the pump will only operate between the min and max displacements, regardless of operating conditions and control signals. This feature will override all other controls: Pressure compensator, load sense, and torque control.

In one example, the minimum displacement feature is accomplished by simply adding a ‘blow hole’ to the control piston that is carefully placed (axial position) along the control piston to relieve control pressure to tank—preventing the pump from de-stroking further. The exact hole position to provide the desired minimum swash angle can be developed and verified through test. The minimum displacement setting is not externally adjustable, but can be changed by removing the control piston and replacing it with a different control piston that has the desired hole location. A maximum displacement setting of the pump can be adjusted by a structure such as an adjustable stop (e.g., a screw stop) that limits a range of movement of the control piston28.

Aspects of the present disclosure can have numerous advantages such as:1. Parallel control—No loss/degradation of pressure compensation or load-sense functions. Hereinafter, pressure compensation and load sense are referred to as base control.a. Much like the parallel nature of pressure compensator and load sense functions, the torque control valve is arranged in parallel to the base control.b. A check valve allows for the base control functions to override the torque control and port pressure/flow to the control piston to further destroke the pump in certain operation conditions.2. High-accuracy control and low cost.a. Hydraulic potentiometer naturally produces continuously variable feedback to enable tracking to the ideal hyperbolic pressure-flow curve.b. Competitive designs utilize a dual-spring arrangement to create an approximation to the required pressure vs. flow hyperbolic curve. This results in a double-hump torque curve with lower accuracy.3. Externally adjustable torque setting—No disassembly required.a. The torque setting is adjusted simply by adjusting the screw in the control valve.b. A single control covers all setting variations (20% to 90% of max.)i. This is possible due to the inherent function of the hydraulic potentiometer in creating a hyperbolic relationship between pressure and displacement.c. Competitive controls require disassembly of the pump to change out spring sets to achieve different torque settings (and maintain accuracy), in addition to adjusting screws.i. A plurality of springs to can be arranged to provide a variable rate spring and approximate the required hyperbola for a given torque setting. However, as the required hyperbola changes for different torque settings the approximation deviates from ideal, creating unacceptable accuracy. Competitive designs are then forced to create several sets of springs to approximate the broad range of required hyperbolic pressure-displacement curves (for the range of torque settings).4. Hydraulic displacement feedback—Reliable and stable performance.a. Since the feedback is purely hydraulic, it provides a smooth and stable feedback, without mechanical failure modes.b. Competitive designs that utilize more complex multi-spring feedback mechanisms are prone to mechanical failure and resonant instabilities.c. Other competitive designs utilize a spring-loaded cam sliding on a surface profile that translates with displacement (usually cut in the control piston). These designs are prone to wear at the cam-slider interface in addition to mechanical failures.5. Available with optional minimum displacement limiter—Ensures a minimal flow (overrides controls).a. Since the minimum displacement limiter is hydraulic, it eliminates the mechanical failure modes.b. Since the limiter is hydraulic, it also provides a ‘soft landing’ rather than a hard stop at the minimum displacement.

The spool66can be configured to have a dual diameter configuration. For example, the spool66can have a larger diameter, A, that is acted upon by Ptc, and a smaller diameter, B, that is acted on by Pout. As described above, an initial pressure drop occurs (“minor losses”) as the oil transitions from the large annulus around the control piston to the small ‘pipe’ as the oil velocity increases. The dual diameter is critical to mitigating the effects of minor losses as well as pump losses.

This pressure drop is proportional to Pout. So the actual Ptc=Pout(1−β)(1−D) where β is a constant coefficient for the fraction of pressure lost due to minor losses.

The area A is sized to compensate for the pressure losses. Ideally, A is sized proportionally larger than B by

AB=11-β
so that the force is balanced across the spool: APtc+Fs=BPout(1−D)

The area difference is further refined to compensate for the inherent mechanical losses in the pump: A piston pump will inherently have losses that vary as a function of pressure and displacement. These inherent losses create an upward or downward trend of the torque limit as a function of pressure/displacement. The area difference is tuned to adjust the trend up/down to further achieve a constant torque limit.

The area difference is further refined to compensate for Poutleakage in the spring cavity: Deviating from ideal, the flow from Poutto Ptcds slightly higher than Ptcto Ptankdue to the leakage from the spring chamber to tank across the spool. This creates slightly higher pressure drop gradient along the first part of the pipe. The area ratio is slightly adjusted to accommodate this leakage.

The deviation from ideal (constant) torque limit due to minor losses, case pressure, and mechanical efficiencies are further mitigated through adjustment of the timing of the hydraulic potentiometer. Recall that in the ideal case:

The actual losses are mitigated further (to get a constant torque limit) by adjusting the axial location of the Ptcsensing location and Poutfeed in the sleeve as well as the start and end of the spiral groove in the piston. This has an effect on both the proportionality and offset of the signal pressure, Ptc, relative to displacement, D.
Ptc=Pout(1−a+bD)

Aspects of the present disclosure can relate to the control piston providing a negatively proportional signal pressure. The negatively proportional signal pressure allows for the mitigating properties of the dual-diameter spool arrangement, as discussed previously, to be designed into the control valve. A positively proportional signal would not allow this—In the positively proportional arrangement (like PVH pump control), Ptcacts directly on the right nose (area A) of the control valve and the spring chamber is at tank pressure. This arrangement also allows better control response to changes in Poutpressure. When the pressure changes in this design, it immediately acts on the nose of the control valve causing the control to react quickly. The positively proportional signal arrangement and direct-acting arrangement (not differential pressure) requires the Poutchanges to be reflected through the spiral groove feature before acting on the nose of the control valve—resulting in a more sluggish response and greater pressure overshoot and slower flow recovery.