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WO2008031483A1 - Système de commande hydraulique pour l'alimentation d'un fluide sous pression régulée en fonction de la demande (régulée en fonction de la détection de la charge) de plusieurs consommateurs hydrauliques - Google Patents

Système de commande hydraulique pour l'alimentation d'un fluide sous pression régulée en fonction de la demande (régulée en fonction de la détection de la charge) de plusieurs consommateurs hydrauliques Download PDF

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Publication number
WO2008031483A1
WO2008031483A1 PCT/EP2007/007163 EP2007007163W WO2008031483A1 WO 2008031483 A1 WO2008031483 A1 WO 2008031483A1 EP 2007007163 W EP2007007163 W EP 2007007163W WO 2008031483 A1 WO2008031483 A1 WO 2008031483A1
Authority
WO
WIPO (PCT)
Prior art keywords
hydraulic
pressure
load
hydraulic pump
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.)
Ceased
Application number
PCT/EP2007/007163
Other languages
German (de)
English (en)
Inventor
Wolfgang Kauss
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Robert Bosch GmbH
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
Priority claimed from DE102007026676A external-priority patent/DE102007026676A1/de
Application filed by Robert Bosch GmbH filed Critical Robert Bosch GmbH
Publication of WO2008031483A1 publication Critical patent/WO2008031483A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

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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
    • F15B11/16—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors
    • F15B11/17—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors using two or more pumps
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
    • 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
    • F15B11/161—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load
    • F15B11/165—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load for adjusting the pump output or bypass in response to demand
    • 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
    • F15B2211/20538—Type of pump constant 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
    • F15B2211/20546—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
    • F15B2211/20576—Systems with pumps with multiple pumps
    • 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
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/265—Control of multiple pressure sources
    • 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
    • F15B2211/50536—Pressure control characterised by the type of pressure control means the pressure control means controlling a pressure upstream of the pressure control means using unloading valves controlling the supply pressure by diverting fluid to the 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
    • F15B2211/605—Load sensing circuits
    • F15B2211/6051—Load sensing circuits having valve means between output member and the load sensing circuit
    • F15B2211/6052—Load sensing circuits having valve means between output member and the load sensing circuit using check 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
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/65—Methods of control of the load sensing pressure
    • F15B2211/654—Methods of control of the load sensing pressure the load sensing pressure being lower than the load pressure

Definitions

  • Hydraulic control arrangement for demand-controlled (pressure-sensing-regulated) pressure medium supply of several hydraulic consumers
  • the invention relates to a hydraulic control arrangement, with the first hydraulic consumers and second hydraulic consumers are provided current-controlled with pressure medium and having the features of the preamble of patent claim 1.
  • a hydraulic control arrangement is used in particular in mobile working machines, such as backhoe loaders.
  • a hydraulic control arrangement with demand flow regulation or according to the load-sensing principle is known from EP 0 566 449 B1.
  • a variable displacement pump is adjusted depending on the highest load pressure of the actuated hydraulic consumers in each case so that the pump pressure is a certain pressure difference above the highest load pressure.
  • the hydraulic consumers, the pressure medium flows through adjustable orifices and these downstream pressure compensators, in the opening direction of the pressure to the respective metering orifice and in
  • Closing direction acted upon by a pending in a rear control chamber control pressure which usually corresponds to the highest load pressure of all supplied by the same hydraulic pump hydraulic consumers. If, during a simultaneous actuation of a plurality of hydraulic consumers, the metering orifices are opened to such an extent that the pressure medium delivered by the stoppered hydraulic pump is less than the total required quantity of pressure medium, the quantities of pressure medium flowing to the individual hydraulic consumers are independent of the respective load pressure of the hydraulic fluid Consumer proportionally reduced.
  • LUDV control load-independent flow distribution
  • a control arrangement with demand flow control can except with a variable with LS pump control valve with a fixed displacement pump and LS-controlled bypass pressure balance, via the pumped by the hydraulic pump and not required by the hydraulic consumers pressure fluid flows back to a tank, can be realized.
  • the losses of unusable energy are higher than when using a variable displacement pump.
  • a hydraulic control arrangement with the features of the preamble of claim 1 is known from EP 1 200 743 B1.
  • the two hydraulic pumps are constant pumps and the two demand flow controllers are bypass pressure balances.
  • the first supply line can be conveyed.
  • second hydraulic pump which is drivable together with the first hydraulic pump, pressure medium in a leading to the second directional control valves, second supply line can be conveyed.
  • first demand flow regulator which is assigned to the first hydraulic pump leads to report the highest load pressure of driven first hydraulic consumers, a first Lastmeldetechnisch.
  • a second demand flow controller is associated with the second hydraulic pump, leading to the message of the highest load pressure of driven second hydraulic consumers a second load reporting line.
  • the second demand flow regulator can also be acted upon by the highest load pressure of the actuated first hydraulic consumers.
  • both the first and the second hydraulic consumers can be supplied with pressure medium both from the first hydraulic pump and also from the second hydraulic pump as well as jointly from both hydraulic pumps , Therefore, in the hydraulic control arrangement according to EP 1 200 743 B1, the second hydraulic pump also conveys via a check valve into the second supply line leading to the second directional control valves.
  • two actuated by an electromagnet 2/2 way valves are available with which the message of the highest load pressure of all simultaneously controlled hydraulic consumers to the first and to the second demand flow regulator can be prevented.
  • the excavator equipment In a backhoe loader, the excavator equipment is normally used when the vehicle is at a standstill. The power of the backhoe loader's internal combustion engine is then only needed for the excavator equipment. By contrast, it is customary to actuate hydraulic loads assigned to the charging equipment while the backhoe loader is being driven. For example, the bucket is raised while driving. Then not only the driving performance but also the power necessary to lift the bucket is to be provided by the engine of the device.
  • the invention has for its object to make a hydraulic control assembly, which is provided for a mobile implement, in particular for a backhoe loader, so that the internal combustion engine of the mobile implement, of which not only the two pumps are driven, but also for driving the working device is used, in its interpretation can be tuned substantially to the mainly driven at a standstill hydraulic consumer.
  • a hydraulic control arrangement with the features of the preamble of claim 1 according to the characterizing part of this claim that upon actuation of a second directional control valve only the second demand flow regulator is subjected to the highest load pressure of the driven second hydraulic consumers. This can ensure that when controlling hydraulic consumers that are also used when driving the mobile implement, always only the second hydraulic pump goes to pressure and loads the drive motor.
  • the first hydraulic pump returns to zero delivery, if it is a variable displacement pump, or returns to the tank via the bypass pressure balance at a low standby pressure, provided that it is a constant displacement pump.
  • each load signaling line is connected directly to the corresponding demand current regulator.
  • the two load-signaling lines are connected to one another via a check valve which opens from the first load-signaling line to the second load-signaling line.
  • a hydraulic steering unit is advantageously supplied with pressure medium by the first hydraulic pump.
  • the hydraulic pumps may be those with a constant displacement, to which a bypass pressure compensator is assigned, or also variable displacement pumps.
  • a flow control valve in particular a flow control valve is provided, which is arranged between the load-sensing line and a drain line to a tank.
  • FIG 1 shows the embodiment of the hydraulic control arrangement according to the invention
  • FIG. 2 shows a longitudinal section through a LUDV directional control valve which can be used in the hydraulic control arrangement according to FIG.
  • the hydraulic control arrangement shown in Figure 1 comprises a first LUDV control block 10 and a second LUDV control block 1 1 in disk construction, each having a plurality of LUDV-way valves 12, as shown in Figure 2, an input disk 13 and 14 and an end plate 15 and 16, respectively.
  • a hydraulic consumer, z. B a double-acting hydraulic cylinder in terms of movement speed and direction of movement are controlled.
  • Each input disk 13 or 14 has a pump connection 17 or 18, a tank connection 19 or 20 and a control connection 21 and 22, respectively. From the connections mentioned, pass channels 23, 24, 25 and 26 which are perpendicular to the disc planes completely through the respective control block 10 and 11 through to the directional valves 12 facing away from ten Design the end plate 15 and 16 extend.
  • the outgoing from the tank port 19 and 20 tank channel 24 is closed by a plug.
  • the channels 25 and 26 are closed in the end plates 15 and 16 by a plug.
  • the channels 23 are fluidly connected to one another by a line 27 running from the one end disk 15 to the other end disk 16.
  • the hydraulic control arrangement has a first hydropump 30 and a second hydraulic pump 31, which are constant-displacement pumps and, in the exemplary embodiment, have the same displacement, ie are the same size.
  • the hydraulic pump 30 discharges pressure medium, which it draws from a tank 33, into a pump line 34, which leads to the pump connection 17 of the control block 10.
  • the second hydraulic pump 31 discharges pressure medium, which also draws it from the tank 33, into a pump line 35, which leads to the pump connection 18 of the control block 11. From the pump connection 17, pressure medium conveyed by the hydraulic pump 30 can be supplied to the directional control valves 12 of the control block via a check valve 36 which is installed in the channel 23 within the input disk 13
  • the tank connections 19 and 20 of the two control blocks 10 and 11 are fluidly connected to the tank 33 via a tank line 39.
  • the input disk 13 contains a first bypass pressure compensator 45 and the input disk 14 a second bypass pressure compensator 46.
  • the two bypass pressure compensators form a first and a second demand current regulator of the control arrangement and are connected with their working connections between the channels 23 and 24 of the control system respective control block is switched.
  • Each input disk is also equipped with a small current regulator 47 which is located between the channel 25 or 26 and the respective channel 24 and limits a pressure medium quantity flowing out of the channel 25 or 26 into the channel 24 for pressure relief of the load signaling line.
  • the control piston of the bypass pressure compensator 45 is acted upon in the closing direction of the pressure compensator by a pressure which prevails in the load-sensing line 25.
  • the control piston of the bypass pressure compensator 46 is acted upon in the closing direction of the pressure compensator by a pressure which prevails in the load-signaling line 26.
  • a pressure compensator acts a compression spring 48 (pressure compensator 45) and a compression spring 49 (pressure compensator 46).
  • the pressure at the pump connection 17 of the control block 10 and in the opening direction of the pressure compensator 46 acts.
  • the end plates 15 and 16 of the two control blocks serve, as already mentioned, in the same way to close the channels 24 and the load reporting lines 25 and 26 and to connect the line 27 to the channels 23 can.
  • a pressure relief valve 50 is housed, which is located between the channel 25 and the tank channel 24 and limits the pressure in the channel 26 to a maximum value.
  • the load reporting lines 25 and 26 associated terminals 21 and 22 of the two control blocks 10 and 11 extends a connecting line 51, in which a of the terminal 22 to the terminal 21 out of the non-return valve 52 is connected.
  • the pressure limiting valve 50 in the control block 11 also limits the pressure in the load-sensing line 25 of the control block 10.
  • the check valve 52 can also be inserted into the input disk 13. be build, between the terminal 21 and the tap of the load pressure for the closing side of the pressure compensator 45th
  • the controllable with the LUDV-way valves 12 hydraulic consumers are referred to as LUDV consumers.
  • the control arrangements shown also include, as further hydraulic consumers, a hydraulic steering system 55, which is indicated in FIG. 1 with a stylized steering wheel.
  • This hydraulic consumer can be supplied solely by the second hydraulic pump 31 with pressure medium.
  • the hydraulic steering 55 is a so-called LS consumer, the pressure medium via a directional control valve not shown in detail, a Zumessblende 57 with variable opening cross-section and an upstream of the metering orifice 57 arranged individual pressure compensator 58 is fed.
  • the steering pressure Downstream of the metering orifice 57, the steering pressure is tapped off and passed via a control line 59 to the control port 22 of the control block 11.
  • a check valve 60 which is arranged with respect to the control terminal 22 parallel to the check valve 52 and blocks from the control port 22 to the steering out.
  • the pressure compensator 58 is acted upon in the opening direction by a compression spring 61 and by the pressure downstream of the metering orifice 57. This pressure is tapped upstream of the check valve 60 on the control line 59. In the closing direction of the pressure compensator 58, the pressure prevailing between this pressure compensator and the metering orifice 57 acts.
  • the compression springs 48, 49 and 61 exert such a force on the respective control piston of the pressure compensators 45, 46 and 58 that, depending on the bias of the compression springs at a lying in the range between 5 bar and 20 bar pressure difference between the two acting on the control piston pressures an equilibrium of forces prevails on the control piston of the pressure compensators.
  • the check valve 60 ensures that the individual pressure compensator 58 of the steering 55 is acted upon by the steering pressure, even if this is not the highest load pressure is. On the other hand, the check valve 60 allows the load pressure of the steering 55 to enter the load reporting passage 26 when the steering has the highest load pressure.
  • Each LUDV directional control valve 12 has a disk-shaped housing 65, which has a lying in the disk plane through the control bore 66 in which a spool 67 is axially movable.
  • the channel 23 passes through an inlet chamber 68 of the control bore 66, which is connectable via control grooves 69 in a collar of the spool 67, regardless of the direction in which the spool is moved from a central position, with an intermediate chamber 70.
  • the opening cross section between the control grooves 69 and the inlet chamber 68 and the intermediate chamber 70 represents a variable metering orifice of the LUDV directional control valves 12 and determines the speed at which a hydraulic consumer is moved.
  • the intermediate chamber 70 has fluid communication with a bore 71, which is also in the plane of the housing 65 and perpendicular to the control bore 66 extends.
  • the control piston 73 of a pressure compensator 74 of which an opening cross-section between the bore 71 and a bridge channel 75 is controlled, in whose two branches in each case a load-holding valve 76 is located from which each branch leads to a further chamber 77 of the control bore 66.
  • Each chamber 77 can be connected via the spool 67 to a consumer control chamber 78, which has a connection to a consumer connection A or B. Beyond each consumer control chamber 78 is still a flow control chamber 79, the connection to the vertically passing through the housing discs 65 channel 24 has.
  • the control piston 73 of the pressure compensator 74 is in the sense of opening the connection between the intermediate chamber 70 and the bridge passage 75 of the Pressure in the intermediate chamber 70, so from the pressure downstream of the metering orifice, which is here designated by the same reference number as the control grooves 69, acted upon.
  • In the closing direction acts on the control piston 73, a pressure which prevails in a control chamber 81 between it and the screw plug 72. Through this control chamber of Lastmeldekanal - here it is the channel 25 - passes.
  • a compression spring 80 acts in the closing direction of the control piston 73, a compression spring 80, which is only slightly biased and the pressure equivalent is, for example, 0.5 bar.
  • the control piston 73 is provided with an axial blind bore 82, which is open to the intermediate chamber 70, and a continuous transverse bore 83, into which the axial bore opens.
  • the inner edge 84 of a shoulder in the bore 71 run over and thereby open a fluid path between the axial bore 82 and the rear control chamber 81 and thus the load reporting channel 25.
  • the fluid path is closed as long as the hydraulic consumer controlled by the directional control valve is not actuated as one and does not have the highest load pressure.
  • the control piston 73 is then in a control position in which it controls an opening cross-section between the intermediate chamber 70 and the bridge passage 75 so that the pressure in the intermediate chamber by the pressure equivalent to the force of the compression spring 80 is higher than the pressure in the rear control chamber 81 If the hydraulic consumer controlled by the directional control valve according to FIG. 2 has the highest load pressure, then the control piston 73 of the pressure compensator 74 completely opens the connection between the intermediate chamber 70 and the bridge passage 75 and now regulates at the control edge 84. It can now act like the control piston a pressure differential valve, which regulates in the rear control chamber 81 a pressure which is lower than the pressure in the intermediate chamber 70 and in the bridge passage 75 upstream of the load-holding valves 76 by the pressure equivalent to the force of the compression spring 80.
  • This pressure is approximately the load pressure of the hydraulic consumer, since the closing spring of the load holding valves used is only weak and also the pressure drop between a control chamber 77 and a control chamber 78 is very small.
  • the load pressure of that hydraulic consumer which has the highest load pressure, respectively via the control piston of one of the pressure compensators 74 contained in the directional control valves 12 in the load reporting channel 25 and thus reported in the rear control chambers on the other pressure compensators.
  • the embodiment according to FIG. 1 is considered in different operating states, whereby it is assumed that the compression springs 48 and 49 are equally biased and both hydraulic pumps 30 and 31 are driven together by an internal combustion engine, for example a diesel engine. If only the hydraulic steering 55 is actuated, the load pressure of the steering via the check valve 60 to the bypass pressure compensator 46, but is not reported because of the check valve 52 to the bypass pressure compensator 45.
  • the pressure compensator 46 now throttles the flow of pressure medium from the pump line 35 to the tank channel 24 of the control block 11 so far that a pressure builds up in the pump line, which is higher than the load pressure of the steering by the pressure equivalent to the force of the compression spring 49.
  • the pressure equivalent may e.g. 20 bar.
  • the individual pressure compensator 58 regulates via the metering orifice a pressure difference corresponding to the pressure equivalent of the pressure spring 61, for example 19 bar.
  • the pressure compensator 45 is completely relieved of its pressure connected to the channel 25 in the input disk 13 control side of pressure. It throttles the flow of pressure medium from the pump line 34 into the tank channel 24 only so much that according to the force of the compression spring 48 in the pump channel 34, a pressure of 20 bar is present.
  • the check valve 36 prevents the pumped by the hydraulic pump 31 pressure fluid flows through the pressure compensator 45 to the tank.
  • the internal combustion engine of the backhoe loader essentially has to apply only the power necessary to drive the hydraulic pump 31 and the drive.
  • the directional control valves 12 of the control block 10 are associated with the excavator equipment of a backhoe loader.
  • the hydraulic consumers of this equipment are usually controlled only when the backhoe loader is, so if the engine does not have to provide power for driving.
  • the directional control valves 12 of the control block 10 is actuated, the highest load pressure all over this way valves 12 simultaneously controlled hydraulic consumers via the load reporting line 25 to the spring side of the pressure compensator 45 and the check valve at the same time to the spring side of the pressure compensator 46 reported.
  • Both pressure compensators throttle the drainage of pressure medium in such a way that a pump pressure which is 20 bar above the highest load pressure is established in the pump lines 34 and 35.
  • the pressure compensator 46 throttles in the channels 23 and 26 a pump pressure which is 19.5 bar higher than the highest load pressure.
  • the pressure compensator 45 remains closed because the pressure equivalent of the spring 48 is 20 bar, but the pressure difference between the highest load pressure and the pump pressure is only 19.5 bar. Promoted by the two hydraulic pumps, but not required pressure fluid flows back via the pressure compensator 46 to the tank. The pressure upstream of the metering orifices is 19.5 bar higher than the highest load pressure.
  • the pressure compensator 45 throttles in the section of the channel 23 of the control block 10 located upstream of the check valve 36 by 19.5 bar and the pressure compensator 46 in FIG the channel 23 of the control block 11 and thus also in the connecting line 27 and located in the downstream of the check valve 36 portion of the channel 23 of the control block 10 to a 20 bar above the highest load pressure pump pressure, as long as the requested pressure medium amount is not the hydraulic pump 35th conveyed pressure medium quantity exceeds.
  • the pressure upstream of the metering orifices is 20 bar higher than the highest load pressure.
  • the pump pressure drops slightly and the pressure compensator 46 closes. Now the pressure compensator 45 throttles a pump pressure which is 19.5 bar higher than the highest load pressure.
  • the pressure upstream of the metering orifices is 19.5 bar higher than the highest load pressure. The small difference in pump pressure from 20 to 19.5 bar hardly makes itself felt.
  • the channel 23 of the control block 11 may be closed in the end plate 16 and the connecting line 27 may be connected directly to the pump line 35, bypassing the control block 11.
  • a privileged pressure medium supply of the steering 55 can be ensured in a conventional manner by a priority valve, with its input to the hydraulic pump 31 and with its primary output to the metering orifice 57 of the steering and with its secondary output to the Terminal 18 of the control block 11 is connected and whose valve spool is acted upon in the sense of a connection of the hydraulic pump 31 with the steering of the pressure in the line 59 and a spring and in the sense of a connection of the hydraulic pump with the port 18 from the pressure at the primary output.
  • the priority valve replaces the individual pressure compensator 58.
  • the priority valve is advantageously integrated in the input disk 14 of the control block 11, which receives an additional connection for the steering system located at the primary outlet of the priority valve. If the space in the input disk 14 is not sufficient for receiving the priority valve and the pressure balance 46, then the pressure compensator can also be accommodated in the end disk 16.
  • the connecting line 27 can be connected, bypassing the control block 11 with the input of the priority valve or directly to the hydraulic pump 31, since the excavator operation, the steering is not needed.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Mining & Mineral Resources (AREA)
  • Civil Engineering (AREA)
  • Structural Engineering (AREA)
  • Fluid-Pressure Circuits (AREA)

Abstract

L'invention concerne un dispositif de commande hydraulique, qui sert à l'alimentation d'un fluide sous pression, régulée en fonction de la demande d'au moins un premier consommateur hydraulique et d'au moins un second consommateur hydraulique, et qui présente des premiers distributeurs à plusieurs voies (12) réglables de manière proportionnelle pour commander le premier consommateur hydraulique et des seconds distributeurs à plusieurs voies réglables de manière proportionnelle pour commander le second consommateur hydraulique. À partir d'une première pompe hydraulique (30), le fluide sous pression peut être transporté par l'intermédiaire d'une soupape de retenue (36) dans une première conduite d'arrivée (23) conduisant aux premiers distributeurs à plusieurs voies. Grâce à une seconde pompe hydraulique (31), qui peut être actionnée conjointement avec la première pompe hydraulique, le fluide sous pression peut être transporté dans une seconde conduite d'arrivée conduisant aux seconds distributeurs à plusieurs voies. Une première conduite de signalisation de charge (25) conduit à un premier régulateur de la demande (45), associé à la première pompe hydraulique, afin de signaler la pression de charge la plus élevée des premiers consommateurs hydrauliques commandés. Une seconde conduite de signalisation de charge (26) conduit à un second régulateur de la demande (46), associé à la seconde pompe hydraulique, afin de signaler la pression de charge la plus élevée des seconds consommateurs hydrauliques commandés. En outre, le second régulateur de la demande peut être également sollicité par la pression de charge la plus élevée du premier consommateur hydraulique actionné. L'invention prévoit que, lors de l'actionnement d'un second distributeur à plusieurs voies, seul le second régulateur de la demande peut être sollicité par la pression de charge la plus élevée du second consommateur hydraulique commandé.
PCT/EP2007/007163 2006-09-13 2007-08-14 Système de commande hydraulique pour l'alimentation d'un fluide sous pression régulée en fonction de la demande (régulée en fonction de la détection de la charge) de plusieurs consommateurs hydrauliques Ceased WO2008031483A1 (fr)

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
DE102006042938.9 2006-09-13
DE102006042938 2006-09-13
DE102007026676.8 2007-06-08
DE102007026676A DE102007026676A1 (de) 2006-09-13 2007-06-08 Hydraulische Steueranordnung zur bedarfstromgeregelten (load-sensing-geregelten) Druckmittelversorgung von mehreren hydraulischen Verbrauchern

Publications (1)

Publication Number Publication Date
WO2008031483A1 true WO2008031483A1 (fr) 2008-03-20

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PCT/EP2007/007163 Ceased WO2008031483A1 (fr) 2006-09-13 2007-08-14 Système de commande hydraulique pour l'alimentation d'un fluide sous pression régulée en fonction de la demande (régulée en fonction de la détection de la charge) de plusieurs consommateurs hydrauliques

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WO (1) WO2008031483A1 (fr)

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0566449A1 (fr) 1992-04-06 1993-10-20 Rexroth-Sigma Distributeur hydraulique combinant la compensation de pression et la sélection de pression maximale
DE19714141A1 (de) 1997-04-05 1998-10-08 Mannesmann Rexroth Ag Hydraulische Steueranordnung
EP1200743A1 (fr) 1999-08-06 2002-05-02 Mannesmann Rexroth GmbH Dispositif de commande hydraulique destine a l'alimentation en agent de pression, reglee par detection de charge, de preference de plusieurs consommateurs hydrauliques
WO2004042235A1 (fr) * 2002-11-07 2004-05-21 Bosch Rexroth Ag Systeme hydraulique a double circuit
WO2004051092A1 (fr) * 2002-11-29 2004-06-17 Bosch Rexroth Ag Systeme hydraulique a deux circuits
US20050072145A1 (en) * 2003-10-01 2005-04-07 Jervis Mark J. Power conserving hydraulic pump bypass compensator circuit
WO2006011836A1 (fr) * 2004-07-28 2006-02-02 Volvo Construction Equipment Holding Sweden Ab Systeme hydraulique et machine de travail comprenant un tel systeme
EP1666735A1 (fr) * 2003-09-26 2006-06-07 Mitsubishi Heavy Industries, Ltd. Dispositif de commande hydraulique pour engin industriel

Patent Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0566449A1 (fr) 1992-04-06 1993-10-20 Rexroth-Sigma Distributeur hydraulique combinant la compensation de pression et la sélection de pression maximale
DE19714141A1 (de) 1997-04-05 1998-10-08 Mannesmann Rexroth Ag Hydraulische Steueranordnung
WO1998045603A1 (fr) * 1997-04-05 1998-10-15 Mannesmann Rexroth Ag Systeme de commande hydraulique
EP1200743A1 (fr) 1999-08-06 2002-05-02 Mannesmann Rexroth GmbH Dispositif de commande hydraulique destine a l'alimentation en agent de pression, reglee par detection de charge, de preference de plusieurs consommateurs hydrauliques
WO2004042235A1 (fr) * 2002-11-07 2004-05-21 Bosch Rexroth Ag Systeme hydraulique a double circuit
WO2004051092A1 (fr) * 2002-11-29 2004-06-17 Bosch Rexroth Ag Systeme hydraulique a deux circuits
EP1666735A1 (fr) * 2003-09-26 2006-06-07 Mitsubishi Heavy Industries, Ltd. Dispositif de commande hydraulique pour engin industriel
US20050072145A1 (en) * 2003-10-01 2005-04-07 Jervis Mark J. Power conserving hydraulic pump bypass compensator circuit
WO2006011836A1 (fr) * 2004-07-28 2006-02-02 Volvo Construction Equipment Holding Sweden Ab Systeme hydraulique et machine de travail comprenant un tel systeme

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