WO2014001043A1 - Dispositif d'entrainement hydraulique - Google Patents

Dispositif d'entrainement hydraulique Download PDF

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Publication number
WO2014001043A1
WO2014001043A1 PCT/EP2013/061564 EP2013061564W WO2014001043A1 WO 2014001043 A1 WO2014001043 A1 WO 2014001043A1 EP 2013061564 W EP2013061564 W EP 2013061564W WO 2014001043 A1 WO2014001043 A1 WO 2014001043A1
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WO
WIPO (PCT)
Prior art keywords
pressure
valve
control
drive according
brake
Prior art date
Application number
PCT/EP2013/061564
Other languages
German (de)
English (en)
Inventor
Peter Loewe
Michael Schuette
Original Assignee
Robert Bosch Gmbh
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Robert Bosch Gmbh filed Critical Robert Bosch Gmbh
Priority to CN201380034063.2A priority Critical patent/CN104379972B/zh
Publication of WO2014001043A1 publication Critical patent/WO2014001043A1/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/4148—Open loop circuits
    • 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
    • F15B11/04—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the speed
    • F15B11/044—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the speed by means in the return line, i.e. "meter out"
    • F15B11/0445—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the speed by means in the return line, i.e. "meter out" with counterbalance valves, e.g. to prevent overrunning or for braking
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/4157—Control of braking, e.g. preventing pump over-speeding when motor acts as a pump
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/46—Automatic regulation in accordance with output requirements
    • F16H61/461—Automatic regulation in accordance with output requirements not involving a variation of the output capacity of the main pumps or motors
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/38—Control of exclusively fluid gearing
    • F16H61/40—Control of exclusively fluid gearing hydrostatic
    • F16H61/46—Automatic regulation in accordance with output requirements
    • F16H61/47—Automatic regulation in accordance with output requirements for achieving a target output speed
    • 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
    • F15B2211/50—Pressure control
    • F15B2211/505—Pressure control characterised by the type of pressure control means
    • F15B2211/50563—Pressure control characterised by the type of pressure control means the pressure control means controlling a differential pressure
    • F15B2211/50581—Pressure control characterised by the type of pressure control means the pressure control means controlling a differential pressure using counterbalance valves
    • 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
    • F15B2211/50—Pressure control
    • F15B2211/52—Pressure control characterised by the type of actuation
    • F15B2211/528—Pressure control characterised by the type of actuation actuated by fluid pressure
    • 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
    • F15B2211/50—Pressure control
    • F15B2211/575—Pilot pressure 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
    • F15B2211/63—Electronic controllers
    • F15B2211/6303—Electronic controllers using input signals
    • F15B2211/633—Electronic controllers using input signals representing a state of the prime mover, e.g. torque or rotational speed
    • 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
    • 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
    • F15B2211/7058—Rotary 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
    • F15B2211/76—Control of force or torque of the output member
    • F15B2211/763—Control of torque of the output member by means of a variable capacity motor, i.e. by a secondary control on the motor
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H59/00—Control inputs to control units of change-speed-, or reversing-gearings for conveying rotary motion
    • F16H59/68—Inputs being a function of gearing status
    • F16H2059/6838—Sensing gearing status of hydrostatic transmissions
    • F16H2059/6876—Sensing gearing status of hydrostatic transmissions the motor speed

Definitions

  • the invention relates to a hydraulic drive according to the preamble of claim 1.
  • a generic hydraulic drive for a winch is known in which a winch drum is driven by a hydraulic motor.
  • the inlet of the hydraulic motor is connected to a pump, preferably a variable displacement pump.
  • the hydraulic fluid flowing from the hydraulic motor flows through the return to a tank T out.
  • a pulling load acts on the winch drum, it can accelerate at insufficient deceleration so that the pumped by the pump pressure medium flow is not sufficient to supply the hydraulic motor with enough pressure medium, so that correspondingly decreases the pressure in the inlet and in unfavorable conditions even cavitation phenomena may occur.
  • a brake valve also called lowering brake valve is provided, which is biased in a closed position and is acted upon in the opening direction by the pressure in the respective inlet.
  • this brake valve is designed so that it is in both directions of rotation of the hydraulic motor, i. is effective for changing supply and return.
  • the brake valve is brought into an open position, so that the flow of hydraulic fluid flowing from the hydraulic motor flows through the brake valve to the tank.
  • the pressure medium volume flow flowing toward the tank is throttled or blocked, so that the hydraulic motor is decelerated accordingly and the pressure in the inlet rises until the brake valve opens again.
  • the invention is therefore an object of the invention to provide a hydraulic drive, prevented in the speed fluctuations due to pressure changes in the inlet or at least reduced.
  • the hydraulic drive according to the invention has a hydraulic motor whose connections can be connected by means of working lines with a pressure medium source and a pressure medium sink and a brake valve assembly, the two working lines are assigned, so that the outflowing pressure medium volume flow can be throttled in both directions.
  • the two effective in the opening direction or in the closing direction control chambers of a brake valve of the brake valve assembly are connected to the input terminals of a selector valve assembly, the larger of the in the control to select adjacent control pressures.
  • the output of the selector valve assembly according to the invention is connected to a pressure relief valve, so that the actuation of the brake valve assembly is co-determined by the pressure set on the pressure relief valve, which is below the maximum expected in each operating condition control pressure.
  • This control pressure usually acts in the opening direction on the brake valve, according to the invention this effective in the opening direction pressure is reduced to the set pressure relief valve pressure, so that even with long supply lines and correspondingly large capacity volumes of the working pressure fluctuations and corresponding speed fluctuations can be reduced.
  • the pressure relief valve is adjusted such that a predetermined maximum speed of the hydraulic motor can not be exceeded and, accordingly, an overspeed is preventable.
  • the pressure limiting valve is made adjustable.
  • This adjustment can be purely electric or electrohydraulic.
  • a spring chamber of the pressure relief valve is internally connected to a tank port of the pressure relief valve.
  • the spring chamber can be acted upon to set the pressure to be limited with a pressure relief control pressure.
  • a variant of the invention provides, in each case to arrange an attenuator in the two lines leading to the above-mentioned control chambers, wherein the tap of the control pressure between the control chamber and the attenuator takes place.
  • the selector valve assembly is a shuttle valve.
  • the selector valve arrangement can also be formed by a check valve arrangement.
  • FIG. 1 shows a circuit diagram of a hydraulic drive according to the invention
  • FIG. 2 shows an enlarged view of a motor unit with brake valve arrangement of the hydraulic drive according to FIG. 1,
  • FIG. 3 shows an enlarged view of a pump unit of the hydraulic drive according to FIG. 1,
  • FIG. 4 shows a partial view of a variant of the embodiment according to FIG. 1
  • FIG. 5 shows a partial view of a further variant of the exemplary embodiment shown in FIG.
  • FIG. 6 shows a diagram for illustrating the braking characteristic of a hydraulic drive according to the invention.
  • the hydraulic drive shown in Figure 1 can be used for example in driving, winch and Turasantrieben.
  • the illustrated hydraulic drive consists of a motor unit 1 and a pump unit 2, which are connected via working lines 4, 6 with each other.
  • the hydraulic capacity and the line resistance are designated by the reference symbols 8, 10.
  • FIG. 2 shows an enlarged view of the motor unit 1 from FIG. 1. Accordingly, this has a hydraulic motor 1, for example, an axial piston adjusting motor whose Swivel cradle via a control cylinder 14 is adjustable.
  • a bottom-side pressure chamber 16 of the actuating cylinder 14 is supplied via a mounted on a housing 18 control unit 20 with control oil.
  • a piston rod-side annular space 22 is connected to the respective pressure side of the hydraulic motor 12, so that the system pressure always acts here and biases the actuating cylinder 14 in the direction of the maximum absorption volume.
  • the bottom-side pressure chamber 16 is connected via the control unit 20 to a tank T.
  • the structure of such control devices is known in principle, so that only the essential components for understanding the invention will be explained.
  • the engine control valve 24 is formed as a continuously adjustable directional control valve, wherein an input port P via a respective check valve 28, 30, both open to the engine control valve 24, with the two working lines 4, 6, which - as explained - to the two working ports the hydraulic motor 12 are connected.
  • By the two parallel check valves 28, 30 is ensured that the higher of the pressures in the working lines 4, 6, the terminal P of the engine control valve 24 is applied.
  • a tank port T of the engine control valve 24 is connected to said tank T, while an output port A is connected to an input port P of the pressure regulating valve 26.
  • the adjustment of the engine control valve 24 by means of a proportional solenoid 32, which is controlled for example via a joystick or the like.
  • the tank port T is connected to the output port A and the port P is shut off.
  • the pressure regulating valve 26 is adjusted by the pressure at the input port P of the pump control valve 24 from its illustrated basic position in the direction of its positions (a). In the spring-biased home position, the input port P of the pressure control valve 26 is connected to an output port B, which in turn is in control oil connection with the bottom pressure chamber 16 of the actuating cylinder 22, so that this relieved in the illustrated basic position of the engine control valve 24 and the pressure control valve 26 to the tank T. is (Vg max).
  • the reduction of the swivel angle and thus the increase of the rotational speed takes place by controlling the proportional magnet 32, wherein the control valve 24 is adjusted in a direction in which a pressure medium connection between the pressure port P and the output port A of the engine control valve 24 is turned on.
  • the pressure control valve 26 initially remains in its illustrated spring-biased home position. Accordingly, then via the check valves 28, 30 of the control chamber in which the higher of the pressures in the working lines 4, 6 acts, connected to the bottom pressure chamber 16 so that flows into this control oil and the actuating cylinder 22 extends due to its area difference and accordingly the
  • Swivel cradle adjusted in the direction of a reduction of the swallow volume (Vg min). The adjustment takes place until an equilibrium of forces between the force of the proportional solenoid 32 and the force of a control spring is reached, which is stretched by the piston of the actuating cylinder.
  • the pressure control via the pressure control valve 26 is superimposed on this electro-proportional control. Increases by a load torque or by reducing the motor pivot angle of the system pressure, which is tapped at the input P of the engine control valve 24 and counter to the force of a control spring 34 acts on the control piston of the pressure control valve 26, a predetermined by the force of the control spring 34 maximum value, the Regulating piston of the pressure control valve 26 is displaced in a direction in which the pressure medium connection B, T is turned on, so that accordingly the bottom-side pressure chamber 16 is relieved to the tank and thus swings the hydraulic motor again. This increase in the absorption volume and the resulting pressure reduction then reduces the control deviation.
  • the hydraulic motor 12 outputs at a constant pressure by increasing the absorption volume from a larger torque.
  • the pressure in the two working lines 4, 6 is each limited by secondary pressure limiting valves 36, 38, each having a non-return function, which allows a flow of pressure medium through the low-pressure secondary Sekundärdruckbegrenzungs- valve in the direction of the then effective high pressure side Sekundärdruckbegrenzungs- valve.
  • a brake valve arrangement 40 is attached to the housing 18. This has a brake valve 42 in a known construction.
  • Two input ports P, P ' are connected to a respective one of the working lines 4, 6, wherein downstream of this connection in each case a check valve 46, 48 is arranged, which allow a pressure fluid flow only in the direction of the hydraulic motor 12.
  • An output port A of the brake valve 42 is connected via a branching line with two drain check valves 47, 49 with downstream of the check valves 46, 48 located portions of the working lines 4, 6.
  • These drain check valves 47, 49 allow only a flow of pressure medium in the direction of the respective working line 4, 6.
  • each attenuator 50, 52 has two parallel damping nozzles 54, 56, the latter a check valve 58 is connected upstream, which allows a control oil flow from the associated damping nozzle 56 and blocks in the opposite direction. Accordingly, the damping nozzle 54 acts to control oil flow toward the respective control chamber, while the damping nozzle 56 determines and damps the control oil flow out of the control chamber.
  • the damping nozzle 54 which determines the inflow to the control chamber, branches off in each case one db line 60, 62 (see also detail 52). These two db lines 60, 62 lead to the input of a shuttle valve 64. Its output is connected to a pressure relief line 66 which is connected to the input of a pressure relief valve 68.
  • this pressure relief valve 68 is designed to be electroproportionally adjustable, so that the maximum pressure to be limited is adjustable. This adjustment takes place via a control unit 70 as a function of the signal of a Drehbaumholz choir 71, via which the speed of the hydraulic motor 12 is detected and the signal is output to the control unit 70.
  • the larger of the control pressures in the control chambers of the brake valve 42 is guided to the pressure relief valve 68 and can then be limited by appropriate adjustment to a maximum value, so that, for example, a maximum speed of the hydraulic motor 12 is not exceeded.
  • the brake valve 42 is adjusted due to the effective in the working lines 4, 6 pressure difference in an opening direction. If the pressure medium supply, for example, via the working line 6 and the pressure fluid discharge via the working line 4, the brake valve 42 is adjusted in the direction of the control positions (a) in which the pressure medium connection from P to A of the brake valve 42 is opened. The pressure medium can then flow via the working line 6 to the hydraulic motor 12 and from the other working port via the upstream of the check valve 48 located part of the working line 4 and the brake valve 42 and then back through the check valve 47 back to the working line 4 and then to the tank T. The speed of the hydraulic motor 12 is constantly monitored.
  • a pressure is set at the pressure limiting valve 68, which is below the voltage applied to the output of the shuttle valve 64 in the regular operation pressure, i. the effective in the opening direction of the brake valve 42 pressure is reduced, so that correspondingly at the motor output a brake pressure is built up, which prevents over-rotation of the hydraulic motor 12.
  • a suitable brake pressure can thus be adjusted in a simple manner as a function of the respective operating state, in order to prevent over-rotation of the hydraulic motor 12.
  • the brake valve assembly 40 is further designed with a brake release valve 72 which is biased by a spring in a locked position and the larger of the pressures in the working lines 4, 6, which in turn is tapped via a shuttle valve 74, in the direction of Opening position is adjustable, in which this greater of the pressures is led to a brake port Br, to which a mechanical brake element is connected, which then depends on From the position of the brake release valve 72 by the higher of the pressures in the working lines 4, 6 is brought out of braking intervention. Since the system pressure applied in the working lines 4, 6 is often incompatible with the pressure required for the brake release, the brake release valve 72 is followed by a brake pressure reducing valve 76, via which the brake pressure is reduced to a brake pressure suitable for the mechanical brake.
  • the pressure medium supply of the hydraulic motor 12 via the pump unit 2 shown in Figure 3. This can be carried out in different ways; in the concrete embodiment, it is an axial piston variable displacement pump unit whose basic structure is known from the prior art, so that only a few components required for understanding are described here.
  • the pump unit 2 has an axial piston variable displacement pump 78, whose delivery volume flow is adjusted by adjusting the pivot angle of a pivoting cradle.
  • the adjustment is made via a control cylinder 80, whose control chamber is connected via a nozzle to the output of a pump control valve 82.
  • This is embodied as an electroproportionally adjustable directional control valve and is controlled via a proportional magnet 84 in the sense of an enlargement of the pivot angle and via an adjustable spring and a control spring 86 arranged in parallel in the sense of a reduction of the
  • Swivel angle minimum displacement applied.
  • said output port A is connected to an input port P, which in turn communicates with an output port of a shuttle valve 88 in fluid communication.
  • the pump pressure in a pump line 90 and on the other hand, a foreign body pressure, which ensures, for example when starting the system that the variable displacement pump 78 is displaced in the direction of its minimum delivery volume (Vg min) at no-current proportional solenoid 84. The larger of these pressures is thus at the port P of the control valve 82.
  • a control line is further connected, which opens via another nozzle in a control chamber of a return cylinder 92, in which a control piston 94 is immersed, which is mechanically connected to the pivoting cradle of the variable.
  • the adjusting movement of the control piston 94 is transmitted via a coupling mechanism 96 to the control spring 86, which is supported on the one hand on this coupling mechanism 96 and on the other hand acts on the control slide of the control valve 82 in the direction of the illustrated basic position.
  • the control piston 94 of the return cylinder 92 is in turn likewise biased by a spring into a basic position, wherein the force acting on the return cylinder 92 and the actuating cylinder 80 spring forces are designed so that the variable displacement in the de-energized non-energized state in the direction of the maximum delivery volume (Vg max ) is biased.
  • Vg max maximum delivery volume
  • a tank port T of the control valve 82 is connected to a port A of a pressure regulating or depressurizing valve 98 whose tank port T is connected to the tank T while a pressure port P is connected to the output port of the shuttle valve 88.
  • the pressure valve 98 is biased by a control spring in its illustrated home position, in which the terminal A is connected to the tank. Accordingly, the tank port T of the pump control valve 82 is connected to the tank T.
  • the continuously adjustable pressure control valve 98 can be adjusted by the pressure in the pump line 90 so that its input port P is connected to the output port A of the pressure control valve 98 and thus the pump pressure in this area is effective.
  • the pressure fluid connection of the control chamber of the actuating cylinder 80 is controlled to the tank T and control oil supplied to the control chamber of the adjusting cylinder 80 so that it extends and correspondingly the variable pump 78 pivots back. That is to say, when the setpoint value set on the pressure regulating valve 98 via its control spring is exceeded, the pump is automatically pivoted back and the control deviation is reduced.
  • This pressure control is superimposed on the above-described EP control.
  • the suction port of the variable displacement pump 78 is connected to the pressure port of a charge pump 100, can be promoted via the pressure medium from the tank T.
  • This charge pump 100 is usually a centrifugal pump, which is effective even at higher speeds.
  • the charge pump 100 facilitates in particular the cold start at low temperatures and correspondingly high viscosity of the pressure medium.
  • a feed pump 102 can be promoted via the pressure medium from the tank T in a feed line 103. This opens in the illustration of Figure 2 between the two flow check valves 47, 49 a.
  • the pump line 90 leads to the connection P of a directional valve 104.
  • This can be adjusted purely electrically or electro-hydraulically or in any other way.
  • the two working ports A, B are connected to the working lines 4, 6.
  • a tank connection T opens in the tank T.
  • the two working connections A, B are connected to the tank T via nozzles in the manner of a floating position.
  • the pressure port P is connected to the working line 6 and the tank port T is connected to the working line 4, which is then located in the discharge line 10.
  • the directional valve 104 is adjusted in the direction "1"
  • the connection P with the working port is reversed B and the tank connection T connected to the working port A.
  • the directional valve 104 may also be designed so that it also connects the pump line to the tank T in the basic position (0). This circulation position is blocked in an adjustment of the directional valve 104 in the direction of "1" or "2".
  • the pressure in the feed line 103 may still be limited via a feed pressure relief valve 106.
  • the pressure regulating valve 98 is still designed with a remote control 108.
  • This consists in principle of a pressure relief valve 1 10 and a running as a 2/2 way valve actuation valve 1 12, the individual or can be used in conjunction.
  • the pressure relief valve 1 10 is made adjustable and limits the pressure in a pressure control line 1 14, which is connected via a nozzle 1 16 to the pump line 90.
  • the pressure in the pressure control line 1 14 acts on the pressure control valve 98 in the direction of its basic position and thus parallel to the pressure control spring.
  • This actuating valve 1 12 can also be designed as a continuously adjustable directional control valve.
  • FIG. 4 shows a variant of the above-described embodiment, in which instead of the electro-proportionally adjustable pressure limiting valve 68, a special design is used in which a spring chamber is internally connected to the tank port of the pressure relief valve 68.
  • the adjustable pressure at the pressure limiting valve 68 which acts on the brake valve 42 in the direction of its (rotational direction-dependent) opening direction, must then be adjusted to the spring of the pressure relief valve 68, wherein the back pressure must be taken into account in the expiring part of the working line.
  • the pressure to be limited is approximately 40 bar plus this back pressure.
  • the flow characteristic of the two check valves 47, 49 (feed) is used as a measuring orifice.
  • Device-specific flow resistances are compensated by means of the inverse shuttle valve 1 18 and thus by the control pressure applied to the control line 16, see above That changes to the devices, which affect the flow resistance of the return lines, have practically no influence.
  • FIG. 5 shows a further variant in which the pressure limiting valve 68 acts in the closing direction on the one hand by the variable force of the pressure limiting spring and additionally by the pressure in a control line 16 which is connected to the outlet port A of an inverse shuttle valve 1 18, both of which Input terminals P, P 'are each connected to one of the working lines 4, 6, wherein the respective lower of the pressures to the terminal A is forwarded.
  • the two pressure control lines 60, 62 via feed check valves 120, 122 connected to each other, wherein the feed line 103 opens into the region between the two feed check valves 120, 122.
  • a conventional pressure relief valve instead of the special valve types in Figures 4 and 5 are used.
  • the brake valve 42 is on the one hand by the pressure difference in the working lines 4, 6, more precisely adjusted by the pressure surplus in the higher pressure working line and the back pressure in the respective sequence in an opening direction, depending on the opening cross-section of the Hydraulic motor 12 running pressure medium flow rate is throttled proportional to this force / pressure difference.
  • the pressure relief valve 68 (in the circuit of Figure 2) driven to adjust the pressure to be limited so that in the opening direction on the Brake valve 42 effective pressure against the actual (with the pressure relief valve 68 closed) acting pressure is lowered, so that due to the reduced opening pressure in the opening cross-section of the brake valve 42 is reduced and accordingly the running pressure medium volume flow is throttled.
  • FIG. This shows a diagram in which the pressure at the engine output, ie the braking torque generating pressure in dependence on the Speed is shown.
  • the pressure relief valve according to the above-described embodiments is set to a maximum pressure of about 70 bar, the maximum speed should be about 1800 revolutions.
  • a hydraulic drive with a brake valve wherein the pressure acting on the brake valve in the opening direction is limited by a pressure limiting valve. This can for example be adjusted depending on the speed or be set to a predetermined pressure.

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

Abstract

Dispositif d'entraînement hydraulique pourvu d'une soupape de freinage, dans lequel la pression qui agit sur la soupape de freinage dans la direction d'ouverture est limitée par l'intermédiaire d'une soupape de limitation de pression. Cette soupape peut être réglée par exemple en fonction de la vitesse de rotation ou également être réglée sur une pression prédéfinie.
PCT/EP2013/061564 2012-06-29 2013-06-05 Dispositif d'entrainement hydraulique WO2014001043A1 (fr)

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DE102012012977.7A DE102012012977B4 (de) 2012-06-29 2012-06-29 Hydraulischer Antrieb
DE102012012977.7 2012-06-29

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US9951861B2 (en) * 2015-03-24 2018-04-24 Ford Global Technologies, Llc Transmission hydraulic control system
DE102015210520A1 (de) * 2015-06-09 2016-12-15 Robert Bosch Gmbh Steuerblock, hydrostatischer Antrieb mit einem Steuerblock und Verfahren für ein Fahrzeug mit einem hydrostatischen Antrieb
CN105485072B (zh) * 2015-12-07 2017-08-25 北京航空航天大学 基于二次调节的舱门瞬态作动装置、系统和控制方法
DE102020208594B4 (de) 2020-07-09 2023-02-09 Robert Bosch Gesellschaft mit beschränkter Haftung Hydrostatischer Zusatzantrieb

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EP0957210A1 (fr) * 1998-05-13 1999-11-17 Poclain Hydraulics Industrie Circuit de fluide hydraulique ayant un moteur hydraulique et un sélecteur à trois positions pour sélectionner la circulation de fluide
EP1475347A2 (fr) * 2003-05-09 2004-11-10 BAUER Maschinen GmbH Ensemble de treuil et procédé pour son fonctionnement
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DE102012012977B4 (de) 2023-06-07
DE102012012977A1 (de) 2014-04-17
CN104379972B (zh) 2016-09-28
CN104379972A (zh) 2015-02-25

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