WO2018039791A1 - Système hydraulique à actionneurs linéaires et à modes hydrostatique et non hydrostatique - Google Patents

Système hydraulique à actionneurs linéaires et à modes hydrostatique et non hydrostatique Download PDF

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Publication number
WO2018039791A1
WO2018039791A1 PCT/CA2017/051019 CA2017051019W WO2018039791A1 WO 2018039791 A1 WO2018039791 A1 WO 2018039791A1 CA 2017051019 W CA2017051019 W CA 2017051019W WO 2018039791 A1 WO2018039791 A1 WO 2018039791A1
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Prior art keywords
pump
hydrostatic
head end
mode
end chamber
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PCT/CA2017/051019
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English (en)
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Travis Kent WIENS
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University Of Saskatchewan
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B11/00—Servomotor systems without provision for follow-up action; Circuits therefor
    • F15B11/02—Systems essentially incorporating special features for controlling the speed or actuating force of an output member
    • F15B11/028—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the actuating force
    • F15B11/036—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the actuating force by means of servomotors having a plurality of working chambers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/205—Systems with pumps
    • F15B2211/2053—Type of pump
    • F15B2211/20546—Type of pump variable capacity
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/205—Systems with pumps
    • F15B2211/2053—Type of pump
    • F15B2211/20561—Type of pump reversible
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/27—Directional control by means of the pressure source
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/315—Directional control characterised by the connections of the valve or valves in the circuit
    • F15B2211/3157—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line
    • F15B2211/31576—Directional control characterised by the connections of the valve or valves in the circuit being connected to a pressure source, an output member and a return line having a single pressure source and a single output member
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/40—Flow control
    • F15B2211/415—Flow control characterised by the connections of the flow control means in the circuit
    • F15B2211/4159—Flow control characterised by the connections of the flow control means in the circuit being connected to a pressure source, an output member and a return line
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/665—Methods of control using electronic components
    • F15B2211/6654—Flow rate control
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/665—Methods of control using electronic components
    • F15B2211/6658—Control using different modes, e.g. four-quadrant-operation, working mode and transportation mode
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/705—Output members, e.g. hydraulic motors or cylinders or control therefor characterised by the type of output members or actuators
    • F15B2211/7051—Linear output members
    • F15B2211/7053—Double-acting output members
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/705—Output members, e.g. hydraulic motors or cylinders or control therefor characterised by the type of output members or actuators
    • F15B2211/7051—Linear output members
    • F15B2211/7055—Linear output members having more than two chambers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders
    • F15B2211/7107—Multiple output members, e.g. multiple hydraulic motors or cylinders the output members being mechanically linked
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders
    • F15B2211/7114—Multiple output members, e.g. multiple hydraulic motors or cylinders with direct connection between the chambers of different actuators
    • F15B2211/7128—Multiple output members, e.g. multiple hydraulic motors or cylinders with direct connection between the chambers of different actuators the chambers being connected in parallel
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/775—Combined control, e.g. control of speed and force for providing a high speed approach stroke with low force followed by a low speed working stroke with high force, e.g. for a hydraulic press
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/785—Compensation of the difference in flow rate in closed fluid circuits using differential actuators

Definitions

  • TITLE A HYDRAULIC SYSTEM WITH LINEAR ACTUATORS AND HYDROSTATIC AND NON-HYDROSTATIC MODES
  • US Patent No. 9, 151 ,018 discloses a hydraulic system that may have a pump with variable-displacement, a first linear actuator, and a second linear actuator coupled to the first linear actuator to operate in tandem.
  • the first and second linear actuators may be connected to the pump in closed- loop manner, and each of the first and second linear actuators may have a first chamber and a second chamber separated by a piston.
  • the hydraulic system may also have an accumulator in fluid communication with the second chamber of only the second linear actuator.
  • US Patent No. 9,051 ,944 discloses a hydraulic unit adapted for connection to master and slave actuator system that includes three valves, the first configured for selective fluid passage between the cap ends, the second configured for selective fluid passage between the slave cap end and an accumulator, and the third fluidly coupled for selective fluid passage between each of a single open circuit pump and the accumulator, and the slave cap end.
  • the valves permit pressurized fluid in the slave cap end to be delivered to accumulator for storage; during extension, the valves permit pressurized fluid from pump and accumulator to be delivered to the slave cap end.
  • a hydraulic system that utilizes two or more linear actuators may be configured so that it can operate in both a hydrostatic mode and a non-hydrostatic mode, and optionally may be switchable between the two modes. This may help provide a system that can be operated in a relatively energy efficient hydrostatic mode during some of its operation, and that can be switched to a more powerful non-hydrostatic mode when required/desired. This may help make the system more energy efficient than a similar system that is operable only in a non- hydrostatic mode.
  • Using linear actuators may make the system suitable for lifting type operations, such as use on heavy equipment and the like.
  • the system may be switchable between modes on the fly, such that an operator need not stop using the system in order to change between modes. This may allow relatively uninterrupted use of the equipment utilizing the hydraulic system, while still allowing for energy savings in the hydrostatic mode when appropriate.
  • a hydraulic system may be switchable between a hydrostatic operating mode and a non-hydrostatic operating mode and may include a first linear actuator having a first rod end chamber, a first head end chamber and a first piston positioned therebetween.
  • a second linear actuator may include a second rod end chamber fluidly connected in parallel with the first rod end chamber, a second head end chamber and a second piston positioned therebetween.
  • the first and second pistons may be movable in unison between extended and retracted positions.
  • a pump may be operable to pump fluid at a first flow rate and a second flow rate.
  • the hydraulic system may be switchable between the hydrostatic mode and the non-hydrostatic mode.
  • the pump may operate at the first flow rate and may be fluidly connected in a closed-loop manner between the first and second rod end chambers and the first head end chamber, the second head end chamber may be fluidly isolated from the first and second head end chambers and the pump, the pump may energize the first head end chamber to urge the first and second pistons toward the extended position with a first force, and the pump may recover energy from fluid driven from the first head end chamber to the first and second rod end chambers when the first and second pistons are urged toward the retracted position by the load.
  • the pump may operate at the second flow rate, the second head end chamber ay be in fluid communication with the first head end chamber, the pump may be fluidly connected between the first and second rod end chambers and the first and second head end chambers, the pump can energize both the first head end chamber and the second head end chamber to urge the first and second pistons toward the extended position with a second force that is greater than the first force.
  • the sum of the piston areas of the first and second rod end chambers may be equal to the piston area of the first head end chamber.
  • a mode selection valve may be fluidly connected between the pump and the second head end chamber.
  • the mode selection valve may be moveable between a first position in which the second head end chamber is fluidly isolated from the hydrostatic pump and the hydraulic system is in the hydrostatic mode, and a second position in which the second head end chamber is in fluid communication with the pump and is connected in parallel with the first head end chamber and the hydraulic system is in the non-hydrostatic mode.
  • the second head end chamber When the mode selection valve is in the first position the second head end chamber may be fluidly connected to a fluid reservoir. When the second piston moves toward the extended position fluid may be supplied to the second head end chamber from the fluid reservoir, and when the second piston moves toward the retracted position fluid may be driven from the second head end chamber into the fluid reservoir. When the mode selection valve is in the second position the second head end chamber may be fluidly isolated from the reservoir.
  • the mode selection valve may travel through at least a first transition position in which fluid communication is established between the second head end chamber, the pump and the fluid reservoir, whereby a leakage portion of the fluid can flow from the pump into the fluid reservoir without passing through the second head end chamber.
  • the second flow rate may be greater than the first flow rate and wherein the pump is operable at a third flow rate that is different than first and second flow rates while the mode selection valve is in the first transition position.
  • the third flow rate may be greater than the first and second flow rates when the hydraulic system is changing from the hydrostatic mode to the non-hydrostatic mode, and may be greater than the first flow rate and less than the second flow rate when the hydraulic system is changing from the non-hydrostatic mode to the hydrostatic mode.
  • the hydraulic system may be switchable between the hydrostatic mode and the non-hydrostatic mode while the first and second pistons are translating and driving the load.
  • the pump may be changeable from the first flow rate to the second flow rate in a pump response time and the mode selection valve may be changeable from the first position to the second position in a valve response time that is equal to or greater than the pump response time.
  • the first and second pistons may translate at a piston speed and the piston speed ay be substantially the same when in both hydrostatic and non- hydrostatic modes.
  • the piston speed may remain substantially constant while the hydraulic system is transitioning between the hydrostatic mode and the non- hydrostatic mode.
  • a controller may be operable to control the pump and the mode selection valve.
  • the controller may monitor a fluid pressure in the first head end chamber, and may be configured to automatically configure the hydraulic system in the hydrostatic mode when the fluid pressure is below a predetermined threshold value, and to change the hydraulic system to the non-hydrostatic mode when the fluid pressure exceeds the threshold value by changing at least one of the operation of the pump and the position of the mode selection valve.
  • the pump may be a bi-directional pump, and may be a variable displacement pump.
  • a method of operating a hydraulic system that is switchable between a hydrostatic operating mode and a non-hydrostatic operating mode and includes a first linear actuator comprising a first rod end chamber, a first head end chamber and a first piston positioned therebetween and a second linear actuator comprising a second head end chamber, a second rod end chamber fluidly connected in parallel with the first rod end chamber and a second piston positioned therebetween that is linked to and movable in unison with the first piston between extended and retracted positions, may include the steps of:
  • the method may include converting the system from the hydrostatic mode to the non-hydrostatic mode while first piston is extending.
  • the method may include operating the pump at a first flow rate when in the hydrostatic mode and operating the pump at a second fluid flow rate that is greater than the first fluid flow rate when in the non-hydrostatic mode whereby the first piston extends at a substantially constant piston translation speed in both the hydrostatic mode and the non-hydrostatic mode.
  • the method may include a mode selection valve fluidly connected to the pump.
  • the second head end chamber may be in fluid communication with the pump when the mode selection valve is in a non-hydrostatic position.
  • the second head end chamber may be fluidly isolated from the pump and may be in fluid communication with a fluid reservoir when the mode selection valve is a hydrostatic position.
  • the method may include simultaneously changing the pump from operating at the first fluid flow rate to operating at the second fluid flow rate and changing the mode selection valve from the hydrostatic position to the non- hydrostatic position.
  • the method may include operating the pump at a third fluid flow rate that is greater than the second fluid flow rate while the mode selection valve is in transition between the hydrostatic and non-hydrostatic positions.
  • the method may include transitioning from the hydrostatic mode to the non-hydrostatic mode while the first piston is extending and without changing the piston translation speed by temporarily positioning the mode selection valve in a transition position as the mode selection valve is moved between the hydrostatic and non-hydrostatic, in which the second head end chamber, pump and fluid reservoir are all in fluid communication with each other.
  • the method may include automatically controlling the operation of the mode selection valve and the pump using a controller.
  • the method may include sensing a fluid pressure in the first head end chamber while the system is operating in the hydrostatic mode, and automatically changing the system to the non-hydrostatic mode when the controller determines that the fluid pressure has exceed a pre-determined transition threshold pressure.
  • the method may include returning the system to the hydrostatic mode when the fluid pressure in the head end chamber falls below the transition threshold pressure.
  • the method may include extracting energy from the system using the pump by retracting the first and second pistons using forces exerted by the load and routing fluid pushed out of the first head end chamber by the retraction of the first piston through the pump to drive the pump in a reverse direction. After passing through the pump to drive the pump in the reverse direction the fluid exiting the pump may be directed into the first and second rod end chambers. Fluid pushed out of the first head end chamber by retracting the first piston may be sufficient to fill both the first and second rod end chambers. Fluid pushed fluid pushed out of the second head end chamber by the retraction of the second piston bypasses the pump.
  • Figure 1 is a schematic representation of one example of a hydraulic system that is operable in a hydrostatic mode and a non-hydrostatic mode, in a hydrostatic configuration;
  • Figure 2 is the system of Figure 1 , in a non-hydrostatic configuration
  • Figure 3 is the system of Figure 1 , in a transition configuration
  • Figure 4 is a schematic representation of another example of a hydraulic system that is operable in a hydrostatic mode and a non-hydrostatic mode, in a hydrostatic configuration;
  • Figure 5 is the system of Figure 4, in a non-hydrostatic configuration
  • Figure 6 is the system of Figure 4, in a transition configuration
  • Figure 7 is a flow chart illustrating one example of a method of operating a hydraulic system that is operable in a hydrostatic mode and a non- hydrostatic mode
  • Figure 8 is a timing diagram illustrating the operation of some components of the hydraulic systems of Figures 1 to 6;
  • Figure 9 is a schematic representation of a hydraulic system used in a simulation
  • Figure 10 is a plot showing orifice curves for a solenoid operated three-way valve, showing an underlapped transition
  • Figure 12 is a plot showing the effect of valve shift time and pump stroke time on the velocity error response as the system shifts from hydrostatic mode to high force mode and back;
  • Figure 13 is a plot showing RMS error with respect to pump response time, for selected valve response times
  • Figure 14 is a plot showing RMS Error with respect to valve response time, for selected pump response times
  • Figure 15 is a plot showing piston translation velocity during mode transition from hydrostatic mode to high force mode
  • Figure 16 is a plot showing piston translation velocity during mode transition from high force mode to hydrostatic mode
  • Figure 18 is a schematic representation of another example of a hydraulic system that is operable in a hydrostatic mode and a non-hydrostatic mode;
  • Figure 19 is a schematic representation of an experimental apparatus based on the hydraulic system of Figure 18;
  • Figure 21 is a plot showing the experimental and modelled pressures for the hydraulic system of Figure 18 in a non-hydrostatic mode, with a 9.3 kN load;
  • Figure 22 is a plot showing the experimental and modelled hydraulic power for the hydraulic system of Figure 18 in hydrostatic mode, with a 9.3 kN load;
  • Figure 24 is a plot showing the experimental and modelled hydraulic efficiency for the hydraulic system of Figure 18 in both the hydrostatic (HE) and non-hydrostatic (HF) modes, with a 9.3 N load;
  • Figure 25 is a plot showing modelled pump and load power as well as losses for a 10 kN external load with the hydraulic system of Figure 18 in a hydrostatic mode;
  • Figure 26 is an enlarged view of the plot of Figure 25 for smaller lifting velocities
  • Figure 27 is a plot showing selected flows for the hydraulic system of Figure 18 in a hydrostatic mode and for a 10kN external load;
  • Figure 28 is a plot showing pressures for the hydraulic system of Figure 18 in a hydrostatic mode and for a 10kN external load;
  • Figure 29 is a plot showing the effect of increasing charge circuit relief valve maximum orifice area on the efficiency of the hydraulic system of Figure 18 in a hydrostatic mode
  • Figure 30 is a plot showing the effect of increasing charge circuit relief valve maximum orifice area the charge circuit pressure for the hydraulic system of Figure 18 in a hydrostatic mode
  • Figure 32 is an efficiency map for full-scale system in non- hydrostatic mode.
  • Figure 34 is a schematic example of another example of a hydraulic system that is operable in a hydrostatic mode and a non-hydrostatic mode;
  • Figure 35 is an efficiency maps for the hydraulic system of Figure 34 in a non-hydrostatic mode with and without a Port B bypass solenoid valve, showing small increase in efficiency for positive velocities (extending) but large improvements for negative velocities (retracting);
  • Figure 36 is a plot showing the effect of cylinder area ratio imbalance on efficiency, for a force of 50 kN;
  • Figure 37 is a simplified schematic for linearized dynamic model
  • Figure 39 is a plot showing pole movement as mass is varied from 1 to 1000 kg, for the model of Figure 37 in non-hydrostatic mode. Hydrostatic mode is similar;
  • Figure 40 is a plot showing the effect of mass on system damping ratio for the model of Figure 37 in non-hydrostatic mode. Hydrostatic mode is similar [0072]
  • Figure 41 is a plot showing the effect of pressure damping on system damping ratio for the model of Figure 37 in non-hydrostatic mode;
  • Figure 42 is a plot showing the effect of pressure damping on pole natural frequency for the model of Figure 37 in non-hydrostatic mode
  • Figure 43 is a plot showing the experimental trace of load force (based on pressures) for a transition of the model of Figure 37 from a hydrostatic to a non-hydrostatic mode time zero;
  • Figure 44 is a plot showing the effect of mass on damping ratio and pressure damping for the model of Figure 37.
  • Some hydrostatic hydraulic systems utilize rotary type actuators and can be configured in a hydrostatic mode while still providing acceptable force/torque.
  • rotary hydraulic systems differ from hydraulic systems that utilize linear/ extending type actuators (i.e. pistons and cylinders) and serve different fluid power needs.
  • linear/ extending type actuators i.e. pistons and cylinders
  • known rotatory hydrostatic systems have not be widely adopted in applications where linear actuators are used - such as to provide lifting/ pushing forces on heavy construction equipment, power diggers, dump trucks, cranes, industrial machines/ presses and the like.
  • a hydraulic system that utilizes linear actuators and that can optionally be operated in a hydrostatic mode to help recover energy and improve the energy efficiency of the system when appropriate, but can also be operated in a relatively higher force, non-hydrostatic mode to help provide increased lifting capacity when needed.
  • a system may be convertible between the different operating modes without changing the system components, and optionally may be changed between operating modes while the system is in use (i.e. on the fly).
  • the system may include any suitable controller, valve and the like so that the transition between hydrostatic and non-hydrostatic operating modes can be done relatively smoothly and optionally without substantially affecting the speed of the linear actuators while extending or retracting.
  • This may help provide a system that is switchable between operating modes while still providing acceptably steady and/or consistent actuator movement.
  • This may be useful if the system is used to lift heavy and/or loose loads (such as an excavator bucket that is full of gravel and/or soil), as sudden changes in the actuator speed may cause the load to shake/ shudder and may contributed to spilling and/or loss of control of the load. Rapid changes in speed and/or acceleration of the actuators may also increase the stresses exerted on the actuators and other system components.
  • the system may include a valve and controller mechanism such that the system can change relatively smoothly between the hydro-static and non-hydrostatic modes without incurring substantial jerks or interruptions in the movement of the actuators, and optionally, while maintaining a generally constant actuator speed. Synchronizing the operation of the valve and pump, and any other components that are used to change system operating modes, may help achieve a sufficiently smooth transition between operating modes.
  • the system may be configured so that when it is in the non-hydrostatic mode its available force is equal to, or at least generally equal to, the lifting force that could be achieved using the same linear actuator in a standard, valve-controlled unbalanced flow system.
  • the systems described herein may be able to achieve such lifting forces and may also be able to provide an actuator that has comparable, and optionally greater maximum actuator speed/velocity.
  • This may also help make the system a modular-type system that can be assembled in various configurations to meet specific needs. It may also simplify service/ maintenance of the system, as individual components may be serviced by generally trained technicians (as opposed to requiring specialized component knowledge), and individual components may be swapped out or replaced using commonly available valves, pumps and the like (without the need to specially order or manufacture a custom replacement part). This may be useful in embodiments where the system is used outside a factory/ shop setting, such as if the system is used on an excavator or other portable piece of equipment. Providing a system that is relatively easily field-servicable using a stock of common replacement parts, possibly even shared between different hydraulic systems across multiple vehicles, may help reduce service time and costs.
  • the system may include a relatively small number of components, and may be free, or at least generally free from flow limiting devices such as throttling valves, throttling orifices which can reduce the efficiency of the system. This may help reduce the energy input required to operate the system.
  • flow limiting devices such as throttling valves, throttling orifices which can reduce the efficiency of the system. This may help reduce the energy input required to operate the system.
  • the system may be arranged to occupy a relatively small physical volume, and preferably to occupy a volume that is not substantially larger than existing, non-hydrostatic linear actuator systems.
  • This may allow the present system to be provided as a retrofit and/or addition to an existing hydraulic system without requiring substantially more space and/or requiring exotic mounting hardware.
  • some embodiments of the system described herein may be generally interchangeable with the current, single linear actuator systems used on the booms of cranes, excavators, power diggers, front end loaders and the like.
  • Systems of the type described herein may be useful in a variety of applications that utilize linear actuators and are utilized in a manner where energy is available for recovery. For example this may include energy recoverable when braking an inertial load or lowering a load against gravity, such as commonly done when using hydraulically powered excavators, backhoes and the like.
  • the hydraulic system may include two linear actuators that are physically linked in some manner, such that both linear actuators move in unison with each other.
  • the linear actuators may be arranged parallel to each other (i.e. such that they extend and retract in unison and in the same direction), opposite each other (such that extension of one actuator corresponds with retraction of the other) or arranged in any other physical configuration that allows for the linking of the actuator movements.
  • suitable actuators are hydraulic cylinders, that include a piston slidable within a housing.
  • FIG. 1 there is provided a simplified schematic of one embodiment of a hydraulic system 100 includes two linear actuators 102 in the form of hydraulic cylinders 102a and 102b.
  • Each cylinder 102 includes a respective piston 104a and 104b, having a piston head 106a and 106b that is slidably received in a corresponding housing 108a and 108b, and a piston rod 1 10a and 1 10b extending from the head.
  • the free ends of the piston rods 1 10a and 1 10b are exposed, and in this embodiment are both connected to a common load 1 12 (for example the boom of an excavator).
  • a common load 1 12 for example the boom of an excavator
  • the pistons 104a and 104b are mechanically linked together such that both pistons 104a and 104b translate in unison with each other. This may be done using any suitable linkage, and in this example, piston rods 1 10a and 1 10b are mechanically linked by the load 1 12 and move simultaneously in the same direction. That is, both pistons 104a and 104b are extended (moving upwardly as shown in Figure 1 ) and retracted (moving downwardly as shown in Figure 1 ) in unison with each other. In this configuration, driving only one of the pistons 104 using pressurized fluid will result in the other one of the pistons translating within its housing, even if it is not separately energized or driven using pressurized fluid.
  • both 102a and 102b define respective head end chambers 1 14a and 1 14b and rod end chambers 1 16a and 1 16b that can be selectably filled/energized with pressurized hydraulic fluid. Pressurizing one or both of the head end chambers 1 14a and 1 14b causes the pistons 104a and 104b to be extended and corresponds to the lifting action in the embodiment of Figures 1 -3.
  • piston 104a may be driven by the hydraulic fluid, and piston 104b may be pulled along with piston 104a due to the linking of the piston rods 1 10a and 1 10b (and vice versa if head end chamber 1 14b were pressurized while head end chamber 1 14a was not).
  • the maximum force that can be exerted by the system 100 on the load 1 12 may be a first force. If both head end chambers 1 14a and 1 14b are simultaneously pressurized, the maximum force that can be exerted by the system 100 on the load 1 12 may be a second force that is greater than the first force, and optionally may be double the first force.
  • Pressuring one or both of the rod end chambers 1 16a and 1 16b can cause the pistons 104a and 104b to retract (i.e. move downwardly as illustrated). If only one of the rod end chambers, such as rod end chamber 1 16a is pressurized, piston 104a may be driven by the hydraulic fluid, and piston 104b may be pulled along with piston 104a due to the linking of the piston rods 1 10a and 1 10b (and vice versa if rod end chamber 1 16b were pressurized while rod end chamber 1 16a was not).
  • the system 100 may be configured so that the piston area of the rod end chamber 1 16a is about 50% of the piston area of the head end chamber 1 14a, and that the piston area of the rod end chamber 1 16b is about 50% of the area of the head end chamber 1 14b. That is, the first piston 104a defines a rod end area facing the rod end chamber 1 16a (i.e. to be acted on by the pressurized fluid in the rod end chamber 1 16a) and an opposing head end area facing the head end chamber 1 14a (i.e. to be acted on by the pressurized fluid in the head end chamber 1 14a).
  • the second piston 104b has an analogous rod end area facing the rod end chamber 1 16b and a head end area facing the head end chamber 1 14b.
  • sum of the first and second rod end areas may be equal to the head end area of the piston 104a or 104b.
  • the sum of the areas of the rod end chambers 1 16a and 1 16b is generally equal to the areas of either one of the head end chambers 1 14a or 1 14b.
  • a closed-loop circuit can be created when operating in hydrostatic mode, that includes the rod end chambers 1 16a and 1 16b, the pump 1 18 and one of the head end chamber 1 14a or 1 14b (head end chamber 1 14a is connected to the hydrostatic loop in the illustrated example).
  • a volume of fluid may be pumped from the head end chamber 1 14a into both rod end chambers 1 16a and 1 16b, and vice versa, without requiring substantial make-up fluid and/or without requiring that excess fluid be dumped from the system in either the extending or retracting phases.
  • a system may include more than two actuators, but may still be configured as a hydrostatic system in a similar way.
  • the actuators may be configured so that the sum of the rod end chamber areas of the actuators is generally equal to the area of one or more of the head end chambers that are to be included in the balanced, closed-loop flow circuit between the rod end chambers and the head end chambers.
  • the piston area of each of the rod end chambers could be A/3 (where A is a total area). If one head end chamber were used in the hydrostatic circuit, its piston area could be A (i.e. the sum of the rod end chambers). If two head end chambers were to be used in the hydrostatic mode, each could have an area of A/2. Other combinations of chambers and areas are also possible.
  • a pump is fluidly connected in the system 100 and is operable to selectably provide pressurized fluid to the head end chambers and rod end chambers.
  • the pump may be any suitable pump including a hydrostatic pump, i.e. any pump that is operable to drive pressurized fluid through the hydraulic circuit and can also be driven in reverse (i.e. in a motor-like mode) by a pressurized fluid flow to allow hydraulic power to be recovered from the hydraulic circuit and converted to mechanical and/or electrical power via the pump.
  • the system 100 includes a pump 1 18 that is a reversible, hydrostatic pump.
  • the pump 1 18 can be driven in a first direction to pressurize the rod end chambers 1 16a and 1 16b, driven in the opposite direction to pressurize one or both of the head end chambers 1 14a and 1 14b (depending on the rest of the system configuration as described herein) a reverse direction.
  • a uni-directional pump and a combination of flow control valves may be used to achieve a similar result.
  • a reversible, hydrostatic pump may be preferable in some embodiments as it may reduce the complexity of the piping/valving required and/or may reduce the overall physical size of the system 100.
  • the same pump is used to energize the cylinders in the lifting phase, and to recover energy from the system when the cylinders are retracting under the load 1 12.
  • the weight of the load 1 12 (whether limited to the structure of the machine, or the structure of the machine plus an additional load such as bucket full of dirt) will tend to act on the cylinders 102a and 102b, and in the illustrated configuration may urge the pistons 104a and 104b toward their retracted positions.
  • the pistons 104a and 104b may be urged toward their retracted positions without having to pressurize either rod end chamber 1 16a or 1 16b, or optionally the rod end chambers 1 16a and 1 16b may be pressurized at a relatively lower pressure than would otherwise be used to retract the pistons 104a and 104b.
  • the forces exerted by the load 1 12 may effectively pump fluid out of the head end chambers 1 14a and 1 14b and through the fluid circuit that includes the pump 1 18. It is this fluid flow that can be utilized for energy recovery in hydrostatic mode.
  • the pump may be a multi- or variable speed pump that is operable to provide at least two different output flow rates to the system 100. This may help provide different fluid flow rates based on the configuration or operating mode of the system. For example the flow rate required to operate the cylinders 102a and 102b at a given speed may be relatively lower when energizing only one of the head end chambers 1 14a or 1 14b, and may be relatively higher when energizing both of the head end chambers 1 14a and 1 14b in unison.
  • the pump may be operable to provide double the flow rate when pressurizing both head end chambers 1 14a and 1 14b, as compared to when pressurizing only one head end chamber, so that the actuator speed can remain substantially constant regardless of which of the chamber(s) 1 14a and 1 14b are being pressurized.
  • This may help make the movements of a machine, such as the boom of an excavator, predictable and generally constant from a user standpoint, while still permitting different lifting forces to be exerted based on the configuration of the system 100.
  • the pump output flow rates may be provided and adjusted using any suitable technique, including by having variable displacement, variable speed or both, such that pressurized fluid can be pumped through the system 100 at least two different flow rates, and optionally at different pressures.
  • a mode selection valve may be positioned in the fluid path between the pump and at least one of the head end chambers (for example 1 14a and 1 14b) and is used to selectively connect or isolate the corresponding head end chamber from the pump and other head end chamber(s). This can allow the targeted head end chamber to be isolated when the system 100 is operating in a hydrostatic mode, and to be fluidly connected to the pump 1 18 to be pressurized when the system 100 is configured in the higher force, non- hydrostatic operating mode.
  • the system 100 includes a mode selection valve 120 that movable between a hydrostatic position ( Figure 1 ), in which fluid connection between the pump 1 18 and the head end chamber 1 14b is interrupted, and a non-hydrostatic position ( Figure 2) in which a fluid connection between the pump 1 18 and the head end chamber 1 14b is restored.
  • Figure 1 a hydrostatic position
  • Figure 2 a non-hydrostatic position
  • the head end chamber 1 14b is fluidly connected to a fluid reservoir, in the form of tank 126.
  • the tank 126 is a non-pressurized tank (i.e. is generally at atmospheric pressure).
  • the pistons 104a and 104b are extended by pressuring head end chamber 1 14a, the corresponding movement of piston 104b causes fluid to be drawn into the head end chamber 1 14b from the tank 126.
  • the pistons 104a and 104b are retracted (either by the load or by energizing the rod end chambers 1 16a and 1 16b, or both), fluid from the head end chamber 1 14b is returned to the tank 126.
  • the mode selection valve 120 may also be positionable in any other suitable positions. While illustrated as a single, two position valve, the mode selection valve may include more than one valve apparatus, may be a multi- position valve, may include a check valve and any other suitable mechanical members that can achieved the desired fluid path configurations.
  • the mode selection valve may be operated using any suitable mechanism and/or control system, including, for example, being manually operable and/or being controlled by a controller.
  • the controller used may be any suitable controller, including, for example a computer, a PLC, a mechanical linkage, a pneumatic or hydraulic control/feedback circuit and the like.
  • the system includes a controller 122 that is configured to the control the operation of the valve 120, by energizing a solenoid 124, to change the system 100 between hydrostatic and non-hydrostatic modes.
  • the system is in hydrostatic mode when the solenoid is not energized, and is changed to the non-hydrostatic mode when the solenoid 124 is triggered by the controller 122.
  • the controller 122 may be configured to automatically trigger the solenoid 124, thereby moving the valve 120, based on one or more input signals 128.
  • the input signal(s) 128 may be based on system conditions such as pressure, piston position, load position, load weight and may automatically adjust the system 100 accordingly. For example, if the system is operating in hydrostatic mode and the system fluid pressure passes a pre-determined threshold (indicating that the load 1 12 may be too heavy to lift using a single cylinder 102), the controller 122 may automatically move the valve 120 to the non-hydrostatic position ( Figure 2), thereby engaging the second head end chamber 1 14b and increasing the available lifting force of the system 100. When such additional lifting force is not required, the controller 122 may automatically shift the valve 120 and return the system to hydrostatic mode.
  • the system 100 may be smoothly switched between modes on the fly: spending time in the relatively more efficient hydrostatic mode when possible, and switching to high-force, non-hydrostatic mode when required.
  • the controller 122 may also control other system components, such as the pump 1 18 and other valves.
  • the controller 122 may control the overall behavior of the system 100 in any suitable manner, including by changing the mode selection valve 120 positions, varying the displacement of the pump 1 18, varying the shaft speed of the pump 1 18, varying a combination of both the displacement and the shaft speed and the like. Such changes may be made automatically based on the performance/ condition of the system, or may be manually selectable by a user, or both.
  • the system may be configured to automatically switch to hydrostatic, energy-recovery mode when the actuators are retracting while loaded, for example when lowering the boom of the excavator.
  • the system 100 may be configured so that it can transition smoothly between operating modes, without causing material changes in the operating speed of the cylinders or otherwise causing jerky or unsteady movement of the cylinders. This may help reduce the chances of a mode change materially impacting the operation of the system 100, and may help facilitate on- the-fly mode changes without causing disruptions to the use of the system 100.
  • the flow rate of the pump 1 18 may be generally doubled to help maintain substantially constant/ steady state translation speed of the piston 104a and 104b when transitioning from the non-hydrostatic mode to the hydrostatic mode, and vice versa.
  • the transition from hydrostatic mode to high-force mode may occur when an external force is about to exceed the force capacity of a single cylinder 102a (i.e. when lifting a heavy load 1 12).
  • head end chamber 1 14a may be at or near the maximum system pressure and/or a pre-set mode change threshold pressure, while head-end chamber 1 14b may be generally at tank pressure.
  • This potentially large pressure imbalance may be at least partially equalized via providing a large flow of fluid through the mode selection valve 120 upon transitioning to high force mode. This may affect the pressures in both the chambers 1 14a and 1 14b and the required pump flow rate to maintain the desired, generally constant piston translation velocity.
  • the time it takes the mode selection valve to transition between the positions of Figures 1 and 2 is understood to define a valve response time
  • the time it takes the pump 1 18 to transition from pumping fluid at a first, steady state flow rate to a second, steady state flow rate is understood to define a pump response time.
  • the mode selection valve 120 can be configured to be positionable in the desirable transition position and may be, for example an underlapped valve, which may allow some leakage flow from supply to tank while the valve is in an intermediate, transition position. This leakage flow may be also be compensated for when supplying and/or releasing fluid from the system 100 during the mode change, as well as when operating the pump 1 18.
  • the mode selection valve 120 may move through the same, or at least an analogous temporary, transition position in which fluid communication between the head end chamber 1 14b and the tank 126 is re-established before fluid communication between the head end chamber 1 14b and the pump 1 18 is fully interrupted.
  • the pump 1 18 output flows during the transition phase may be different than either of the steady state flow rates used in the hydrostatic or non-hydrostatic modes.
  • the pump output flow rate may be dynamically adjusted, optionally via the controller 122, to help account for the varying flow conditions and dynamics.
  • the pump flow rate may be temporarily increased to a third flow rate that is greater than either of the steady state hydrostatic or non-hydrostatic flow rates to provide a temporary boost in the flow.
  • the pump flow rate may temporarily be set at an intermediate flow rate that is between the steady state hydrostatic or non-hydrostatic flow rates, or optionally may be below the hydrostatic steady state flow rate.
  • the pump 1 18 may be configured to have a sufficiently fast pump response time such that these temporary flow rates may be utilized while the mode selection valve 120 is in transition, and optionally so that the pump 1 1 8 is returned to operating at the appropriate steady state flow rate by the time the mode selection valve 120 transition is complete.
  • FIG. 8 The relative operations of the pump 1 18 and mode selection valve 120 of one possible embodiment of the system 100, as the system transitions from hydrostatic mode to non-hydrostatic mode is illustrated in Figure 8.
  • the pump flow rate as a function of time, is illustrated by line 130 and the flow rate of fluid flowing between the tank 126 and head end chamber 1 14b is illustrated by line 132.
  • the position of the valve 120 is illustrated in the middle portion of Figure 8, with the available percentage of flow area in communication with the tank 126 represented by line 134, and the percentage of flow area in communication with head end chamber 1 14b represented by line 136.
  • the upper portion of Figure 8 illustrates the system pressure as recorded at the head end chamber 1 14a using line 138, and pressure at the head end chamber 1 14b represented by line 140.
  • [001 14] In some embodiments of the system 100, when the system 100 is operated in hydrostatic mode, with the pump 1 18 flow 130 is a first steady state flow rate (represented by line 142 in Figure 8) while lifting a load.
  • the system pressure in the head end chamber 1 14a is monitored by the controller 122. As the forces exerted by the load 1 12 increases, the system pressure may reach a pre-determined transition pressure level 144 at a given time, represented by line 146. This may trigger the controller 122 to activate the mode selection valve 120 at the first time 146, to change the system 100 to non-hydrostatic mode, to help provide additional lifting force.
  • the mode selection valve 120 can be actuated, and will reach its second position (the position of Figure 2) by second time 148.
  • the lateral spacing between lines 146 and 148 can represent the valve transition time 150.
  • the flow area 1 34 in communication with the tank 126 is reduced from about 100% to about 0%, while the flow area 136 providing communication between the pump 1 18 and the head end chamber 1 14b increases from about 0% to about 1 00%, and during the transition time 1 50 the relevant flow areas 134 and 136 are both non-0% at the same time (representing the desired transition flows permitted by an underlapping valve).
  • the pump flow rate 130 may be increased to account for the transition flow characteristics.
  • the pump flow rate 130 may temporarily reach a third, transition flow rate level (line 1 52), before settling into the second steady state flow level (line 154).
  • the transition flow rate 152 is greater than the non-hydrostatic steady state flow rate 154, and the hydrostatic flow rate 142.
  • the pump 1 18 has a relatively fast pump transition time (represented by distance 156) and responds quickly enough so that the pump flow rate 130 has returned to the non-hydrostatic steady state flow rate 1 54 when the mode selection valve 120 reaches its non-hydrostatic position at time 148.
  • FIG. 1 -3 The schematic diagrams provided in Figures 1 -3 are intended to illustrate the general arrangement of the components and the flow paths that contribute to the desired operation of the system 100. However, the schematics are somewhat simplified and do not show all of the specific components that may be present in a specific hydraulic system. For example, standard components such as relief valves, anti-cavitation valves, charging systems, cooling systems and the like are well understood by those skilled in the art, and have been excluded from the present schematics for clarity.
  • Figures 4-6 are schematic representations of how the elements of the system 100 may be incorporated into a more realistic, real-word hydraulic circuit, and include various other hydraulic components such as an input/output shaft 160 connected to the pump 1 18, a charge pump 162, filter 164, optionally pump displacement controller 166 (which may by communicably linked to controller 122 and/or may be integral with the controller 122), charge relief valve 168, port relief valves 170 and make-up check valves 172.
  • the elements of the system 100 operate in the same manner as described herein, when in the hydrostatic configuration ( Figure 4), the high force, non-hydrostatic mode ( Figure 5) or when in the temporary transition configuration (Figure 6).
  • FIG. 7 one example of a method 200 of operating a hydraulic system that is switchable between hydrostatic and non-hydrostatic modes (such as system 100) using a controller (such as controller 122) is illustrated.
  • the method 200 may begin with engaging the controller at step 202 and optionally, reading a desired operator selected piston translation speed 204.
  • the controller 122 may also sense the pressure in the head end chamber 1 14a at step 206.
  • the controller 122 can query if the system is currently configured in hydrostatic mode, non-hydrostatic mode or is transitioning between modes.
  • the controller 122 may then check if the pressure in head end chamber 1 14a is below a pre-determined transition pressure level at step 210. If not, the controller 122 may signal the pump 1 1 8 to continue operating at the hydraulic mode steady state flow rate at step 212, and repeat steps 202-208 on a pre-predetermined basis. If the pressure sensed in step 210 is greater than the pre-determined transition pressure level , the controller 122 may proceed to step 216 and begin changing the system 100 to hydrostatic mode.
  • the controller 122 can trigger the mode selection valve 120 to change positions. While the valve 120 is in transition, the controller 122 may measure/sense the flow rate of fluid flowing to the tank 126 at step 220 (for example as a result of the leakage during the valve transition) and may determine suitable adjustments to the pump flow rate to compensate for any such tank flows.
  • the controller 122 may then measure the pressures in the head end chambers 1 14a and 1 14b in step 222, and optionally may determine suitable adjustments to the pump flow rate to help compensate for compressibility of the fluid under the sensed operating conditions.
  • the controller 122 can check to determine if the system pressure is still at a level that justifies operating in non-hydrostatic mode at step 226. If so, the controller 122 can command the pump 1 18 to operate at the non-hydrostatic mode, steady state flow rate. If not, at step 230 the controller 122 may begin switching the system 100 to hydrostatic mode by triggering the valve at step 232. While the valve 120 is in transition, the controller 122 may measure/sense the flow rate of fluid flowing to the tank 126 at step 220 and may determine suitable adjustments to the pump flow rate to compensate for any such tank flows.
  • the controller 122 may then measure the pressures in the head end chambers 1 14a and 1 14b in step 222, and optionally may determine suitable adjustments to the pump flow rate to help compensate for compressibility of the fluid under the sensed operating conditions.
  • the controller 122 may determine the desired pump out put level, for example based on the information obtained in steps 220 and 222, and may send a corresponding command signal to the pump 1 18 (or associated sub-controller) to adjust the pump flow rate. The method may then return to step 202, and may be repeated at any desired frequency.
  • the controller 122 may measure/sense the flow rate of fluid flowing to the tank 126 at step 220 and may determine suitable adjustments to the pump flow rate to compensate for any such tank flows. [00129] The controller 122 may then measure the pressures in the head end chambers 1 14a and 1 14b in step 222, and optionally may determine suitable adjustments to the pump flow rate to help compensate for compressibility of the fluid under the sensed operating conditions.
  • the controller 122 may determine the desired resultant pump flow rate level, for example based on the information obtained in steps 220 and 222, and may send a corresponding command signal to the pump 1 18 (or associated sub-controller) to adjust the pump flow rate. The method may then return to step 202, and may be repeated at any desired frequency.
  • a dynamic model 1 100 of the system was constructed to help study the dynamics during mode switching events.
  • the model layout was based on a schematic illustrated in Figure 9, which incorporates representations of the components of system 100, as well as other hydraulic components. Elements corresponding to features of the system 100 are labelled using like reference characters, indexed by 1000.
  • the modelled system 1 100 includes a pump 1 1 18 and solenoid dynamics, static pressure and anti-cavitation valves, and cylinder models 1 1 02a and 1 102b, including compressible volumes 1 1 14a, 1 1 14b, 1 1 16a and 1 1 16b, and a mode selection valve 1 120.
  • P A i and P A2 are the pressures in the head end chambers 1 1 14a and 1 1 14b
  • P B is the pressure in both rod end chambers 1 1 16a and 1 1 16b
  • a A and A B are the piston areas of the head and rod end of each piston 1 104.
  • the compressible volumes in the three chambers are assumed to not cavitate and to have constant bulk modulus ⁇ with volumes that vary with the piston stroke:
  • V A o and V B o are the cylinder dead volumes and x max is the cylinder stroke.
  • C d is the discharge coefficient
  • A is the orifice area
  • ⁇ is the pressure drop
  • v and p are the fluid kinematic viscosity and density
  • D is the orifice's hydraulic diameter
  • P CR is the transition pressure, related to the critical Reynolds Number, Re cn by [00138]
  • Q A ⁇ A2 ( ⁇ - PA2, A MA2 ) and from tank to the second cylinder
  • Q TA 2 f(Pr - PA2, A T A2)-
  • valve orifices are assumed to have linear orifice curves shown in Figure 10 which shows the orifice curves for the solenoid operated three-way valve, showing underlapped transition.
  • Valve dynamics are approximated by a rate limiter that allows the valve to fully shift from one position to the other in time shift-
  • Relief valves and check valves are both assumed to respond infinitely fast with flow calculated using the same orifice model as in equation 7.
  • the orifice area of each valve is assumed to increase linearly from the cracking pressure, P ck , to the maximum orifice area, A MAXL over the pressure override range, P or :
  • Figures 12-14 quantify the effect of the response speed of the pump and valve. This demonstrates that a fast pump response is critical to low mode transition error. It also demonstrates the somewhat unusual situation where a fast solenoid valve can cause increased error, if the pump cannot respond fast enough to keep up.
  • An experimental test was also performed to help verify the stability of the control law developed above. Two standard cylinders were connected to a Vickers PVB5 variable displacement axial piston pump modified with direct displacement control. This pump has a fast displacement response, with a rise time between 15 and 35 ms. A Parker D3W solenoid valve was used to control the transition between hydrostatic and high-force modes, with a rated shift time of 25- 35 ms.
  • the ideal control law in Eqn. 21 was implemented using a NIDAQ data acquisition and control system connected to a computer running Matlab Realtime Windows Target. The solenoid was also controlled by this system.
  • the system was instrumented with drag-type flow meters to measure Q s and QTA2, and with pressure transducers on P M , P A 2 and P B .
  • the cylinder's velocity was estimated by measuring the cylinder outlet flow using a calibrated orifice. This orifice also simulated a load on the cylinder. Cavitation was avoided via an elevated tank pressure of 2.4 MPa.
  • Figures 15 and 16 show the results of transition events (with time set to zero when the controller initially commands the solenoid to shift).
  • the controller gain was varied between 0.5 and 1 .0. With the gain set to 1 .0, the response was oscillatory and had a steady state error when reenergizing the solenoid. The best gain in this case is near 0.8, as demonstrated by the root mean squared error in the velocity, shown in Figure 17.
  • the amount of energy that may potentially be recovered by this system is dependent on the work cycle of the particular hydraulic circuit.
  • One sample work cycle was used to provide one quantified example of the potential for energy recovery over a typical loading cycle.
  • Other work cycles could be analyzed in an analogous manner. This analysis estimates the fractions of time that the system spends in each mode and the ideal energy that can be recovered in each.
  • the example work cycle selected was from a 40-ton hydraulic excavator loading a truck. An experienced operator was asked to load a series of trucks located on the same level as the excavator, at a swing angle of 90°. The soil was sandy loam. Repositioning via tracks and waiting for truck positioning were not considered part of the work cycle and these sections were removed from the data.
  • Table 2 shows the breakdown of the total energy available for each function in this example.
  • the swing circuit shows the highest fraction of energy recoverable (defined by the fraction of energy recoverable to work done), and many manufacturers have commercial examples available that apply hydrostatic systems, electrical swing drives or accumulator systems to recover this inertial energy.
  • the maximum absolute value of energy recoverable is in the boom circuit, demonstrating that significant savings may be available if this energy can be recovered.
  • Table 2 Energy breakdown for an excavator truck-loading cycle.
  • a schematic example of another hydraulic system 2100 includes two linear actuators 2102 in the form of hydraulic cylinders 2102a and 2102b that are configured as single-rod cylinders, each with a 2: 1 piston area ratio.
  • Each cylinder 2102 includes a respective piston 2104a and 2104b, having a piston head 2106a and 2106b that is slidably received in a corresponding housing 2108a and 2108b, and a piston rod 21 10a and 21 10b extending from the head.
  • the free ends of the piston rods 21 10a and 21 10b are exposed, and in this embodiment are both connected to a common load 21 12.
  • the pistons 2104a and 2104b are mechanically linked together such that both pistons 2104a and 104b translate in unison with each other.
  • piston rods 21 10a and 21 10b are mechanically linked by the load 21 12 and move simultaneously in the same direction. That is, both pistons 2104a and 2104b are extended (moving upwardly as shown in Figure 1 ) and retracted (moving downwardly as shown in Figure 1 ) in unison with each other. In this configuration, driving only one of the pistons 2104 using pressurized fluid will result in the other one of the pistons translating within its housing, even if it is not separately energized or driven using pressurized fluid.
  • both 2102a and 2102b define respective head end chambers 21 14a and 21 14b and rod end chambers 21 16a and 21 16b that can be selectably filled/energized with pressurized hydraulic fluid.
  • the system 2100 includes a pump 21 18 that is a reversible, hydrostatic pump.
  • the pump 21 18 can be driven in a first direction to pressurize the rod end chambers 21 16a and 21 16b, driven in the opposite direction to pressurize one or both of the head end chambers 21 14a and 21 14b (depending on the rest of the system configuration as described herein) a reverse direction.
  • the pump 21 18 may have any of the attributes or features described in relation to pump 1 18 herein.
  • the hydraulic system 2100 can be operated in a hydrostatic, and relatively high efficiency mode, in which one of the hydrostatic pump's 21 18 ports are connected to one head-end chamber 21 14a and 21 14b and the other port is connected to both rod-end chambers 21 16a and 21 16b.
  • the other head-end chamber 21 14a or 21 14b is connected to low pressure (in this case supplied by a charge circuit).
  • the single head-end flow is approximately equal to the two rod-end flows, so the flows are generally balanced.
  • This mode exhibits about half the maximum extending force of a conventional system with similarly sized cylinders, but may have about double the maximum extending velocity and may provide the ability to recover energy to the shaft when braking or lowering a load.
  • the system 2100 can also include various other hydraulic components such as an input/output shaft 2160 connected to the pump 21 1 8, a charge pump 2162, a filter, an optional pump displacement controller (which may by communicably linked to controller 122 and/or may be integral with the controller 122), charge relief valve 2168, port relief valves 2170 and make-up check valves 2172.
  • the elements of the system 21 00 operate in the same manner as described herein, when in the hydrostatic configuration , the high force, non-hydrostatic mode or when in the temporary transition configuration.
  • the hydraulic system 2100 includes a mode selection apparatus that can be used to change the system between its hydrostatic mode, in which fluid connection between the pump 21 18 and the head end chamber 21 14b is interrupted, and its non-hydrostatic position in which a fluid connection between the pump 21 18 and the head end chamber 21 14b is restored.
  • the system 1 00 includes a single valve 120 that can function as the mode selection apparatus
  • the hydraulic system 2100 includes a hydrostatic mode valve 2120a and a non-hydrostatic mode valve 2120b.
  • the valves 2120a and 2120b may be independently operable and may be provided with respective solenoids that can be controlled by the controller 2122 (or any other suitable controller).
  • ⁇ is the shaft angular speed (rad/s) and D is the pump's displacement (m 3 /rev).
  • the pump's 21 18 internal and external leakage are assumed to be laminar:
  • R AB , R AT , and fl BT are the effective resistances of each leakage path.
  • the charge pump 2162 is assumed to be an ideal fixed displacement pump:
  • QAC Qor(A RVAC , P A — P ) — Qor (A CVAC , P c — P A ) (29) where A RVAC and A CVAC are defined as in eq 28.
  • K RVB is an area-related coefficient
  • a tanh function is used to smooth the infinite derivative near zero area, with the extent of the distortion controlled by P REF .
  • a CVAC follows the linear area relationship in eq 28.
  • the cylinder(s) 2102 is modelled as an ideal cylinder with no leakage (cylinder leakage is lumped with the pump's R AB although it is small relative to pump leakage). The flow into each head end is
  • F c is the ideal Coulomb friction force
  • B f is the viscous damping coefficient
  • tanh is used instead of a sign function to avoid discontinuities around zero velocity.
  • the velocity v ref is used to control what is considered a "small" velocity around zero. The net force on the cylinder is then
  • a CA2 is the HE valve's 2120a orifice area, while in High Force (non- hydrostatic) mode:
  • a A1A2 is the HF valve's 2120b area.
  • Qsnet Q B - Qs - QBC + QAB ⁇ QBT (42)
  • Q Alnet -Q A1 - Q AB - Q A1C + Qs - QAIA2 - QAT (43)
  • Qcnet Qc + QBC + QAIC ⁇ QcT ⁇ QcA2 (44)
  • the above equations may constitute a nonlinear system of equations that may be solved to determine the steady state operating point. This was achieved by using the Matlab "fsolve" function to numerically solve for values of P A , P B , P c , and v that force the flow and force continuity equations 36, 42, 43, and 44 to zero. Although it is possible for this set of equations to result in more than one solution, the present simulation did not find any situations where this is case.
  • a performance metric used here is the hydraulic efficiency, defined as
  • the apparatus 2300 included a Hydogear model PY pump (not shown), which is a hydrostatic pump designed for use on lawn equipment, with a main pump maximum displacement of 21 .8 cc/rev and a 4.1 cc/rev charge pump.
  • the apparatus 2300 also included a frame 2302, an arm 2304 movably coupled to the frame 2302 that can be loaded with weights 2308 (to simulate loads 1 12, 21 12) and a pair of cylinders 2306 to simulate the actuators 2102a and 2102b (with only one cylinder 2306 schematically shown in Figure 19).
  • the distance 2310 was about 608mm
  • the distance 2312 was about 183mm
  • the distance 2314 was about 81 1 mm.
  • the pump in this experiment integrates workport check and relief valves as well as charge pump relief valve with experimentally determined parameters found in Table 3.
  • a check valve was included to avoid cavitation at the charge pump outlet, but it is not believed that this valve ever opened during the experiments conducted.
  • the pump was run by a 20 HP electric motor at 1750 rpm. This motor may be oversized for this application, which may help ensure a constant shaft speed.
  • valve 2120a In order to help facilitate adjustment of the workport B relief pressure, an external relief valve was installed, Hydraforce model RV08-22.
  • the experimentally-determined parameters are found in Table 3.
  • the solenoid valves used to control the modes were Parker model DSL102C for the HF valve and a Hydraforce SF20-22 for the HE valve.
  • the HE valve (such as valve 2120a) may preferably be selected to be larger than the HF valve (such as valve 2120b) to help avoid the possibility of cavitation, it is recognized the valve used in the apparatus 2300 may be substantially oversized.
  • Solenoid valves were controlled over a J 1939 bus via Hydraforce EVDR 201 A valve drivers.
  • the pump's swash plate position was controlled using a rotary hydraulic actuator with a separate power supply, operated by a servo valve controlled using an analog proportional closed loop controller.
  • a load was provided by the apparatus shown in Figure 19. This allows for a gravitational and inertial loading, by applying up to 247 kg of weight, which can apply up to 1 3.8 kN of force to the cylinders 2306.
  • This apparatus 2300 was designed to mimic the nonlinear force characteristics of a front-end loader or other such piece of heavy equiment, in smaller scale.
  • the cylinders' 2306 position was measured using a MTI Instruments model LTC-300-200-SA laser displacement transducer, providing a calibrated analog output. Pressures were measured using STW model M01 -CAN J1939 pressure transmitters. Transmitters with a range of 25 MPa were used for P A , P AZ , and P B , while P c was measured using a range of 7 MPa. J 1939 signals were acquired using a Vector CANboard XL interfaces, while analog signals were acquired using a National Instruments PCIe-6251 interface. All data was logged at a 1 0 ms sample rate.
  • Figures 20 and 21 show the effect of cylinder velocity on pressures in HE (hydrostatic) and HF (non-hydrostatic) mode, with both model and experimental values. These data were recorded with the apparatus 2300 loaded with 204 kg of weights, which corresponds to approximately 9.3 kN force at the cylinders when the boom is horizontal.
  • Figures 22 and 23 show the main pump, charge pump, and load hydraulic power (i.e. pressure times flow in each case). The load flows are calculated from load velocity and the pump nominal flow is calculated from swash plate angle. Therefore, the hydraulic system efficiency calculated from this data (shown in Figure 24), includes the effect of leakage (volumetric efficiency), but not mechanical frictional losses in the pump or load (mechanical efficiency).
  • Figures 25 and 26 are plots comparing the modelled power and losses in the system.
  • the inventors note that the leakage losses are approximately constant with load velocity, which dominate the efficiency at low flows.
  • the inventors believe that a better efficiency curve can be expected by either selecting a smaller pump with less leakage or using more flow by increasing the cylinder size (also increasing the maximum force) when implementing the systems 100, 2100 and the like.
  • the actuators may not be perfectly balanced.
  • m v P A A A - P B A B — BfV— F (49)
  • a A and A B represents the effective piston areas (both cylinders for A B and one cylinder for A A in HE mode and both in HF mode).
  • the load mass can vary considerably while the system is in use. This mass is a parameter that has an effect on the dynamic response and any variability of the mass can have an effect on performance of the system.
  • Figure 39 shows the pole movement as the load mass (e.g. load 512, or 1 12, 21 12, etc.) is varied.
  • the system may, in some configurations, be underdamped while the damping ratio decreases quickly as the mass increases (damping ratio is shown in Figure 40). This may cause a different character for the system depending on load, which was also evident in the behavior of the experimental apparatus.
  • the system could utilize the measured velocity at predetermined points in the fluid circuit to add a damping term to the pump controller (such as controller 122 or 2122).
  • a damping term such as controller 122 or 2122.
  • the pump controller such as controller 122 or 2122
  • cylinder position sensors may not be commonly installed in the type of equipment that may utilize the system 100, 2100 in a retrofit capacity, and differentiating the signal may produce a relatively very noisy result.
  • the inventor has developed and implemented a pressure-based term in the pump controller to help achieve some desired system damping. This may help facilitate a relatively a low-cost and relatively low-noise solution as compared to some alternative damping apparatus.
  • the pump flow is modified to include the pressure rate as
  • Equation 38 through 44 can be modified by replacing each with + K.
  • the effect of the pressure damping gain is shown in Figures 41 and 42.
  • the damping ratio of the conjugate poles may be increased, which may help reduce oscillations.
  • the effects of this damping may be generally traded off against a slower dynamic response and a less stiff system.
  • a system for a given application may be designed with a pre-determine, and acceptable combination of these features/values.
  • This transition creates a disturbance in the system, which may cause the base system to bounce for some time.
  • the pressure damping system described herein may help reduce this bouncing. While it may not entirely eliminate the oscillations, perhaps due to only damping oscillations in P A1 , while ignoring P B , it may help provide a real and/or perceived improvement in the controllability of the system from the system user's perspective.
  • a variable mass can affect the tuning of the pressure damping gain.
  • the gain required to bring the damping ratio to unity varies by about an order of magnitude as the mass is varied from 100 to 1000 kg.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Fluid-Pressure Circuits (AREA)

Abstract

L'invention concerne un système hydraulique comprenant deux actionneurs hydrauliques linéaires ou plus et pouvant fonctionner en mode de fonctionnement hydrostatique et en mode de fonctionnement non hydrostatique, et pouvant commuter entre ceux-ci.
PCT/CA2017/051019 2016-08-30 2017-08-30 Système hydraulique à actionneurs linéaires et à modes hydrostatique et non hydrostatique WO2018039791A1 (fr)

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Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP7400915B1 (ja) 2022-09-27 2023-12-19 いすゞ自動車株式会社 ポンプシステム及びそれを備えた車両

Citations (3)

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Publication number Priority date Publication date Assignee Title
US7634911B2 (en) * 2007-06-29 2009-12-22 Caterpillar Inc. Energy recovery system
US20130098012A1 (en) * 2011-10-21 2013-04-25 Patrick Opdenbosch Meterless hydraulic system having multi-circuit recuperation
US9151018B2 (en) * 2011-09-30 2015-10-06 Caterpillar Inc. Closed-loop hydraulic system having energy recovery

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US7634911B2 (en) * 2007-06-29 2009-12-22 Caterpillar Inc. Energy recovery system
US9151018B2 (en) * 2011-09-30 2015-10-06 Caterpillar Inc. Closed-loop hydraulic system having energy recovery
US20130098012A1 (en) * 2011-10-21 2013-04-25 Patrick Opdenbosch Meterless hydraulic system having multi-circuit recuperation

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP7400915B1 (ja) 2022-09-27 2023-12-19 いすゞ自動車株式会社 ポンプシステム及びそれを備えた車両
JP2024047728A (ja) * 2022-09-27 2024-04-08 いすゞ自動車株式会社 ポンプシステム及びそれを備えた車両

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