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EP1092095B2 - Circuit hydraulique - Google Patents
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EP1092095B2 - Circuit hydraulique - Google Patents

Circuit hydraulique Download PDF

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Publication number
EP1092095B2
EP1092095B2 EP99936360A EP99936360A EP1092095B2 EP 1092095 B2 EP1092095 B2 EP 1092095B2 EP 99936360 A EP99936360 A EP 99936360A EP 99936360 A EP99936360 A EP 99936360A EP 1092095 B2 EP1092095 B2 EP 1092095B2
Authority
EP
European Patent Office
Prior art keywords
pressure
consumer
load
bypass channel
control
Prior art date
Legal status (The legal status 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 status listed.)
Expired - Lifetime
Application number
EP99936360A
Other languages
German (de)
English (en)
Other versions
EP1092095B1 (fr
EP1092095A1 (fr
Inventor
Thomas Weickert
Erich Adlon
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Bosch Rexroth AG
Original Assignee
Bosch Rexroth AG
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Filing date
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Application filed by Bosch Rexroth AG filed Critical Bosch Rexroth AG
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Publication of EP1092095B1 publication Critical patent/EP1092095B1/fr
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Classifications

    • 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
    • F15B11/003—Systems with load-holding 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
    • 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
    • F15B11/05—Systems essentially incorporating special features for controlling the speed or actuating force of an output member for controlling the speed specially adapted to maintain constant speed, e.g. pressure-compensated, load-responsive
    • 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
    • F15B11/16—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors
    • 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
    • F15B2211/20553—Type of pump variable capacity with pilot circuit, e.g. for controlling a swash plate
    • 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
    • F15B2211/25—Pressure control functions
    • F15B2211/253—Pressure margin control, e.g. pump pressure in relation to load 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
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
    • F15B2211/3053—In combination with a pressure compensating valve
    • F15B2211/3054—In combination with a pressure compensating valve the pressure compensating valve is arranged between directional control valve and 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
    • F15B2211/30—Directional control
    • F15B2211/31—Directional control characterised by the positions of the valve element
    • F15B2211/3144—Directional control characterised by the positions of the valve element the positions being continuously variable, e.g. as realised by proportional 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
    • F15B2211/30—Directional control
    • F15B2211/32—Directional control characterised by the type of actuation
    • F15B2211/329—Directional 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
    • F15B2211/505—Pressure control characterised by the type of pressure control means
    • F15B2211/50509—Pressure control characterised by the type of pressure control means the pressure control means controlling a pressure upstream of the pressure control means
    • F15B2211/50545—Pressure control characterised by the type of pressure control means the pressure control means controlling a pressure upstream of the pressure control means using braking valves to maintain a back 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
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/605—Load sensing circuits
    • F15B2211/6051—Load sensing circuits having valve means between output member and the load sensing circuit
    • F15B2211/6054—Load sensing circuits having valve means between output member and the load sensing circuit using shuttle 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
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders
    • 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
    • F15B2211/78—Control of multiple 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
    • F15B2211/80—Other types of control related to particular problems or conditions
    • F15B2211/88—Control measures for saving energy

Definitions

  • the invention relates to a hydraulic circuit for controlling at least one lower-load and higher-load consumer according to the preamble of patent claim 1.
  • Such circuits are used inter alia for controlling mobile machines, such as excavators.
  • Such circuits are used inter alia for controlling mobile machines, such as excavators.
  • Such a load-sensing circuit is known, for example, from EP 0 566 449 AS.
  • This circuit has a variable displacement pump, which can be controlled so that it produces at its output a pressure which is higher than the highest load pressure of the hydraulic consumers by a certain amount.
  • a load-sensing controller is provided, which can be acted upon by the pump pressure in the direction of reducing the stroke volume and the highest pressure at the consumers and by a compression spring in the direction of increasing the stroke volume.
  • the setting in the variable displacement difference between the pump pressure and the highest load pressure corresponds to the force of the aforementioned compression spring.
  • Each of the consumers is associated with an adjustable orifice plate with a downstream pressure compensator, via which the pressure drop across the orifice plate is kept constant, so that the hydraulic fluid flowing to the respective consumer depends on the opening cross section of the orifice and not on the load pressure of the consumer or on the pump pressure.
  • the pressure compensators of all actuated hydraulic consumers are adjusted in the closing direction, so that all hydraulic fluid flows to the individual. Consumers are reduced by the same proportion. That is, with downstream pressure compensator, the flow rates to the consumers are always in proportion to the opening cross-sections of the orifices. Due to this load-independent flow distribution (LUDV), all driven consumers move at a speed reduced by the same amount.
  • LUDV load-independent flow distribution
  • variable displacement pump is usually equipped with a pressure control and with a power control, via which the maximum possible pump pressure or the maximum deliverable by the variable displacement power (excavator performance) are adjustable. These pressure and power controls are superimposed on the load-sensing control.
  • WO95 / 32364 discloses a control arrangement via which, when a limit load pressure is exceeded, only the load pressure of the lower-load hydraulic consumer is reported to the load-sensing regulator of the variable displacement pump.
  • This limit load pressure is chosen so that the supply of the other hydraulic consumer is guaranteed.
  • this is achieved by the spring chamber of the pressure compensator of the lower-load consumer via a pressure limiting valve assembly is connected to the tank.
  • the pressure relief valve opens the connection to the tank, so that the spring chamber of the pressure balance of the lower-load consumer relieved and the control piston is brought into its open position, in which the load pressure of this consumer is reported in the load pressure signaling line.
  • a disadvantage of this control arrangement is that a partial volume flow is discharged to the tank and thus can not be used for the consumer control. The efficiency of this control is therefore comparatively low. Another disadvantage is that generated by the return of the hydraulic fluid to the tank heat in the system and thus pump power is destroyed.
  • the present invention seeks to provide a control arrangement by which a sufficient supply of all consumers is guaranteed with minimal device complexity.
  • a particularly simple circuit is obtained when the pressure compensator upstream measuring orifice is formed by a proportional directional control valve, wherein the bypass channel in response to the valve spool position of the proportional directional control valve is alsêtbar. Due to the control of the bypass channel dependent on the control of the proportional valve, the individual pressure compensator only acts in the fine control range, in which comparatively small hydraulic fluid volume flows through the pressure compensator.
  • the structure can be further simplified if the bypass channel is formed in the valve spool of the proportional directional control valve and is controllable by a control edge of the valve spool bore.
  • a check valve arrangement is provided therein.
  • two working connections of a consumer are controlled via the proportional valve.
  • the bypass channel is assigned to only one of the working ports, so that, for example, in the lifting function, a flow through the bypass.
  • bypass channel is opened only after a certain stroke of the proportional valve, so that no bypass flow is formed at the beginning of the control.
  • the valve spool of the proportional directional valve is preferably formed with a central speed part and two outer direction parts, which are each associated with a terminal of the consumer.
  • the bypass channel extends within the valve spool from Geschwindkeitsteil towards the direction part, so that the pressure compensator is bypassed.
  • the pressure loss in the bypass channel can be minimized if it opens with oblique and radial bores in the outer circumference of the valve spool.
  • a part of a circuit diagram for a hydraulic circuit for controlling a mobile implement, such as an excavator is shown.
  • This excavator has several consumers, such as a boom, a spoon, a handle, a chassis drive and a slewing drive, which are supplied by a variable displacement pump 2 with hydraulic fluid.
  • a cylinder 4 for actuating a spoon and a cylinder 6 for actuating the excavator boom are shown as a consumer schematically.
  • the hydraulic fluid delivered by the variable displacement pump is led via a pump line 12 with branch lines 12a, 12b to the two consumers 4 and 6, respectively.
  • an adjustable orifice plate 14a, 14b is formed in each branch of the pump line 12 (12a, 12b. As will be explained in more detail below, these orifices 14a, 14b are designed as speed parts of a proportional valve.
  • a pressure compensator 16a, 16b Downstream of each metering orifice 14a, 14b, a pressure compensator 16a, 16b is connected in each case.
  • the control piston of this 2-way pressure compensator is acted upon in the opening direction via a control line 18 with the pressure downstream of the orifice 14a, 14b and in the closing direction via a load control line 20 with the highest load pressure, which is tapped from a load pressure signaling line 22. About this the highest load pressure is also led to the load-sensing controller 8.
  • the volume flow of the load at the lower end of the load decreases to a value which is predetermined by the maximum pump power. A large part of the power is destroyed in the regulating pressure compensator of this consumer.
  • the load-lower consumer b is assigned a bypass channel 32 in the control illustrated in FIG. 1, which allows a bypass of the pressure compensator 16a.
  • the bypass channel 32 branches off downstream of the orifice plate 14a and opens into the working line 24a to the consumer 6.
  • a suitable control device 34 is provided which shuts off the bypass channel 32 in the basic position and aufêtt depending on the opening cross section of the orifice 14a.
  • the switching device provided with the reference numeral 34 may be any device which is suitable for shutting off the bypass channel 32 and réelle juryn in response to the control of the orifice plate 14a.
  • FIG. 2 the circuit diagram of a valve disc 35 of a valve block for realizing the circuit shown in Fig. 1 is shown.
  • the valve disc 35 includes the pressure compensator 16a, a proportional valve 36, through the speed part of the orifice plate 14a is formed and the bypass channel 32, and the other, described in more detail below connection lines of the hydraulic elements.
  • the proportional valve 36 in addition to the orifice 14 a and a directional part for controlling the load A, B, and the control of the bypass channel 32 are integrated.
  • the proportional valve 36 has a pump port P, two working ports A, B, which are connected to the cylinder chambers of a differential cylinder b or with a hydraulic motor. Furthermore, an output port P1 to the pressure compensator 16a, a bypass port U, two input ports R, S of the directional part and a tank port T are formed on the proportional valve 36.
  • valve spool 38 of the proportional valve 36 are biased by two compression springs 41a, 41b in their basic position. In this basic position, ports P, A, B, U and S are shut off while ports P1 and R are connected to the tank.
  • valve spool 38 The end faces of the valve spool 38 are subjected to control pressures P ST , so that it can be moved out of its spring-biased basic position.
  • the output port P1 is connected via the pump line 12a to the input port Q of the pressure compensator 16a.
  • the control line 18 via which the pressure downstream of the orifice 14a (proportional valve 36) is reported to the left in Fig. 2 end face of the pressure compensator 16a.
  • the load pressure of the consumer 6 is connected via the load-sensing line 20 to the load pressure signaling line 22 and guided to the spring side of the pressure compensator 16 a.
  • the output terminal C of the pressure compensator 16a is connected via lines 40, 42 to the input terminals R and S of the directional part. In the lines 40, 42 there are two check valves 56a, 56b, which prevent a backflow of the hydraulic fluid from the direction part to the pressure compensator 16a.
  • the tank connection T is connected via a tank line 44 to the tank.
  • the pressure compensator 16a By the pressure compensator 16a, the pressure drop across the metering orifice 14a is kept constant and independent of the load pressure when the proportional valve 36 is actuated, so that the volume flow to the consumer 6 is proportional to the opening cross section of the metering orifice 14a.
  • valve spool 38 When applying a control pressure P ST, for example, to the left end face of the proportional valve 36, the valve spool 38 is shifted to the right, so that the orifice plate 14a for connecting the terminals P, P1 is turned on. In the fine control range, that is to say in the first part of the valve spool stroke, the connection to the bypass duct connection U is still blocked.
  • the hydraulic fluid is supplied via the working line 12a to the input terminal Q and via the control line 18 to the left end side of the control piston of the pressure compensator 16a, so that it is moved to its control position for keeping constant the pressure drop across the orifice plate 14a.
  • the thus adjusted hydraulic fluid flow is then passed via the line 40, the terminals R, A to the working port of the consumer 6, while via the working port B and the tank line 44, the hydraulic fluid from the consumer 6 is fed back to the tank.
  • the connection S is closed.
  • the bypass channel 32 is opened by the valve spool 38, so that the hydraulic fluid directly into the Line 40 flows.
  • the volume flow to the pressure compensator 16a is reduced or even completely shut off, so that a larger volume flow to the consumer 6 is performed. This increase in the volume flow also leads to a drop in the system pressure when the load-bearing consumer 4 is driven to the stop.
  • Fig. 3 shows a section through a directional control valve segment, by which the circuit shown in Fig. 2 is realized.
  • the directional control valve segment has a valve plate 52, in which receiving bores for the valve spool 38, the pressure compensator 16a, two pressure relief valves 54a, 54b and the two non-return or load-holding valves 56a, 56b are formed.
  • the valve plate 52 In the valve plate 52 are further provided the two working ports A, B, two control terminals 58a, 58b for controlling the proportional valve 36, a pump port P, at least one connection for the load pressure signaling line 22 and a tank connection.
  • the valve spool 38 has in its central region a control collar 60, which forms the orifice plate 14a in cooperation with a web 62 of the valve bore.
  • the valve slide 38 is biased by the two compression springs 41a, 41b into its basic position, in which no flow through the metering orifice 14a takes place.
  • the control of the proportional valve 36 is effected by applying a control pressure to the two control terminals 58a and 58b, which are connected via control lines to the spring chamber 64a and 64b of the proportional valve 36.
  • a nozzle is formed with a check valve, through which a damping of the valve spool movement is possible.
  • the control collar 60 is provided in the region of its end faces with a plurality of control notches 64 and 66, via the pressure medium from a connected to the pump port P annulus 68 can be performed to the input terminal Q, so that the lower in Fig. 3 end face of the control piston 72nd the pressure compensator 16a with the pressure downstream of the orifice can be acted upon.
  • the orifice plate 14a is formed by cooperation of the control notches 64 with the one control edge of the web 62, while with a shift to the left the control notches 66 open the connection from the annulus 68 to the pressure compensator 16a out ,
  • the input port Q of the pressure compensator 16a is formed as an axial connection, so that the fluid pressure also acts on the lower end face 70 of the control piston 72.
  • the output terminal C is formed as a radial connection and opens into the lines 40 and 42. In these lines 40, 42, the load-holding valves 56a, 56b are arranged, which prevent backflow from the valve spool 38 to the pressure compensator 16a out and allow a flow in the reverse direction.
  • each working port A, B is associated with a directional part, via a working port A and B with a Line 40, 42 or with the tank T is connectable.
  • the right-trained in Fig. 3 direction part for the terminal B has three axially spaced control covenants 74, 76 and 78.
  • the control covenants 76 and 78 are each provided with control notches 80 and 82, which arranged to the between these control cords 76, 78 Open radially recessed section.
  • the working part A associated direction part of the valve spool 38 is formed only by two spaced control covenants 84, 86.
  • control notches 88 are formed, which correspond in function to the control notches 80 of the control collar 78.
  • a plurality of circumferentially distributed oblique holes 90 which are connected to a common axial bore 92. This passes through the control collar 8 to the left end portion of the valve spool 38.
  • the end stop 94 of the valve spool is screwed into the axial bore 92, so that the left end portion is closed.
  • valve slide 38 shows a detailed representation of the valve slide 38 in the middle region of this axial bore 92.
  • a retention valve is provided in the axial bore 92, the valve body 96 is biased by a compression spring 97 against a valve seat 98.
  • a radial hole star 100 and an oblique hole star 102 Downstream of the valve body 96 open a radial hole star 100 and an oblique hole star 102.
  • the radial hole star 100 is blocked by a web 104 of the receiving bore 103 of the valve spool 38.
  • the oblique bore star 102 opens into the radially recessed portion between the control collars 84 and 86.
  • the biased against the valve seat 98 valve body 96 prevents hydraulic fluid from the port A can flow into the axial bore 92. A flow in the opposite direction is practically not prevented because the compression spring 97 is weak.
  • the geometry of the radial bore star 100 and the oblique bore star 102 is selected such that upon a displacement of the valve spool 38 to the left over these stars 100, 102, the connection from the working port A to the tank port T can be opened.
  • control notches in the right end face area of the control collar 84 are used.
  • valve spool 38 is moved in the illustration of FIG. 3 to the right, so that the control notches 64 in conjunction with the web 62 control the connection from the pump port P to the input port Q of the pressure compensator on.
  • the upper end face 105 of the control piston 72 lying at the top in FIG. 3 is acted upon by the force of a control spring 106 and by the load pressure, which is tapped by a circumferential groove 110 via a control edge and an angular bore 108 in the control piston 72.
  • a control spring 106 By the pressure applied to the input terminal Q downstream of the metering orifice 14a of the control piston 72 is deflected upward and the output port C open until an equilibrium of forces on the control piston 72 is established.
  • the load holding valve 56a is opened and the hydraulic fluid via the line 40 and the control collar 86 with the control notches 88 to the working port A out.
  • connection between the working port B and the tank port T is opened up via the control collar 76 and the control notches 82 assigned to the working port B, so that the hydraulic fluid can flow back from the consumer into the tank.
  • the oblique bores 90 of the bypass channel 32 are not yet opened by the control edge 107.
  • control edge 107 controls the bypass channel 82, so that the hydraulic fluid or at least a partial volume flow to the working port A is performed.
  • the system pressure drops, so that the load lower consumer 6 can be operated at a higher speed.
  • bypass channel 32 is assigned only to the working port A, which is required for the lifting function of the consumer.
  • the other working port B a further bypass channel can be assigned, which would then have an identical structure as the above-described working port.
  • a load-lower consumer for example, a boom is controlled.
  • the bypass channel 32 is opened in the manner described above, so that the hydraulic fluid flow Q increases to the lower-load consumer (dashed line). Due to this increase of the hydraulic fluid volume flow to the lower-load consumer, the pressure from the system pressure p SYS drops to a lower level p *.
  • the pressure level p * can be adjusted so that the pressure drops, for example, from a pressure of 240 bar to a pressure p * of 200 bar.
  • bypass channel 32 is integrated into the proportional valve 36.
  • bypass channel is realized via external circuits.
  • a LUDV circuit for controlling at least one lastniedrigeren and a higher-load load, each consumer an orifice plate and a downstream pressure compensator are assigned to keep the pressure drop across the orifice constant.
  • the pressure compensator of the lower-load consumer is assigned an openable bypass channel, via which the pressure compensator of this consumer can be bypassed.

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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 circuit à répartition du flux indépendante de la charge, servant à commander au moins un consommateur à charge basse (4) et un consommateur à charge élevée (6), un diaphragme de mesure (14a, 14b) et une soupape de maintien de pression montée en aval (16a, 16b) , servant à maintenir constante la réduction de pression sur le diaphragme de mesure, qui sont associés à chaque consommateur (4, 6). A la soupape de maintien de pression (16a) du consommateur à charge basse est associé un canal de dérivation (32) pouvant être ouvert, par lequel la soupape de maintien de pression de ce consommateur peut être évitée.

Claims (9)

  1. Circuit hydraulique pour activer au moins un consommateur de charge moins élevée et un consommateur de charge plus élevée (4, 6), avec une pompe à cylindrée variable à puissance régulée (2), dont le réglage peut être modifié en fonction de la pression de charge la plus élevée du consommateur (4, 6), dans lequel est prévu, entre la pompe à cylindrée variable (2) et chaque consommateur (4, 6), un orifice de mesure (14a, 14b) réglable avec une balance manométrique montée en aval (16a, 16b), dont le piston de réglage (72) peut être alimenté dans la direction de fermeture par des pressions de charge prélevées par les consommateurs (4, 6) et dans la direction d'ouverture par la pression en aval de l'orifice de mesure (14a, 14b), caractérisé en ce que la sortie (P1) de l'orifice de mesure (14a) associé au consommateur de charge moins élevée (6) peut être reliée, via un canal de dérivation (32), en contournant la balance manométrique individuelle (16a) montée en aval de l'orifice de mesure (14a), avec au moins un raccord de service (A) pour le consommateur de charge moins élevée (6), pour que lors de l'ouverture commandée du canal de dérivation (32), la pression du système baisse à un niveau plus bas que lorsque le canal de dérivation (32) est fermé et que le courant de fluide hydraulique augmente en direction du consommateur de charge moins élevée.
  2. Circuit hydraulique selon la revendication 1, caractérisé en ce que l'orifice de mesure (14a, 14b) est formé à travers une soupape proportionnelle (36), pouvant être raccordé à un raccord de pompe (P) ou à un réservoir (T) via le raccord de service (A, B), et en ce que le canal de dérivation (32) est pilotable en fonction de la position du robinet-vanne de la soupape proportionnelle (36).
  3. Circuit hydraulique selon la revendication 2, caractérisé en ce que le canal de dérivation (32) est formé dans le robinet-vanne de soupape (38) et en ce qu'il est pilotable par une rampe hélicoïdale de la soupape proportionnelle (36).
  4. Circuit hydraulique selon l'une quelconque des revendications précédentes, caractérisé en ce que dans une soupape de retenue (96, 97, 98), qui empêche un écoulement de liquide hydraulique du consommateur (6) vers l'orifice de mesure (14a), est disposé le canal de dérivation (32)
  5. Circuit hydraulique selon l'une quelconque des revendications 2 à 4, caractérisé en ce que la soupape proportionnelle (36) présente deux raccords de service (A, B) pour le consommateur (6), et en ce qu'un canal de dérivation (32)est disposé au niveau de chaque raccord de service (A, B).
  6. Circuit hydraulique selon l'une quelconque des revendications 2 à 5, caractérisé en ce que le canal de dérivation (32) est piloté après une course prédéterminée du robinet-vanne de soupape (36).
  7. Circuit hydraulique selon l'une quelconque des revendications 2 à 6, caractérisé en ce que le robinet-vanne de soupape (36) présente une zone de vitesse formant l'orifice de mesure (14a) et agencée de manière à peu près centrale et deux parties directionnelles, par lesquelles le liquide hydraulique peut être conduit du raccord de sortie (Q) de la balance manométrique (16a) vers un raccord de service (A ; B) ou de l'autre raccord de service (B ; A) vers un raccord de réservoir (T), de sorte que le canal de dérivation (32) s'étend de la zone de vitesse vers les parties directionnelles.
  8. Circuit hydraulique selon l'une quelconque des revendications 2 à 7, caractérisé en ce que le canal de dérivation (32) débouche d'une part sur des alésages inclinés (90) dans la zone de vitesse et d'autre part, sur un alésage en étoile radial (100) et/ou un alésage en étoile incliné, (102) en aval de la soupape de retenue (96, 97, 98) dans la zone d'une partie directionnelle.
  9. Circuit hydraulique selon l'une quelconque des revendications précédentes, caractérisé en ce que la pompe réglable (2) est réglée en pression.
EP99936360A 1998-06-29 1999-05-31 Circuit hydraulique Expired - Lifetime EP1092095B2 (fr)

Applications Claiming Priority (3)

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DE19828963A DE19828963A1 (de) 1998-06-29 1998-06-29 Hydraulische Schaltung
DE19828963 1998-06-29
PCT/DE1999/001591 WO2000000747A1 (fr) 1998-06-29 1999-05-31 Circuit hydraulique

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EP1092095A1 EP1092095A1 (fr) 2001-04-18
EP1092095B1 EP1092095B1 (fr) 2003-03-26
EP1092095B2 true EP1092095B2 (fr) 2007-04-18

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JP (1) JP4520041B2 (fr)
KR (1) KR100636863B1 (fr)
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WO (1) WO2000000747A1 (fr)

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KR20010071687A (ko) 2001-07-31
JP4520041B2 (ja) 2010-08-04
EP1092095B1 (fr) 2003-03-26
DE59904746D1 (de) 2003-04-30
DE19828963A1 (de) 1999-12-30
KR100636863B1 (ko) 2006-10-19
JP2002519596A (ja) 2002-07-02
EP1092095A1 (fr) 2001-04-18
US6367365B1 (en) 2002-04-09
WO2000000747A1 (fr) 2000-01-06

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