EP1149246A1 - Dispositif de commande pour au moins deux consommateurs hydrauliques, et soupape a pression differentielle correspondante - Google Patents

Dispositif de commande pour au moins deux consommateurs hydrauliques, et soupape a pression differentielle correspondante

Info

Publication number
EP1149246A1
EP1149246A1 EP00906203A EP00906203A EP1149246A1 EP 1149246 A1 EP1149246 A1 EP 1149246A1 EP 00906203 A EP00906203 A EP 00906203A EP 00906203 A EP00906203 A EP 00906203A EP 1149246 A1 EP1149246 A1 EP 1149246A1
Authority
EP
European Patent Office
Prior art keywords
pressure
control
valve
inlet
connection
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.)
Granted
Application number
EP00906203A
Other languages
German (de)
English (en)
Other versions
EP1149246B1 (fr
Inventor
Martin OBERHÄUSSER
Thomas Weickert
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.)
Bosch Rexroth AG
Original Assignee
Mannesmann Rexroth AG
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 Mannesmann Rexroth AG filed Critical Mannesmann Rexroth AG
Publication of EP1149246A1 publication Critical patent/EP1149246A1/fr
Application granted granted Critical
Publication of EP1149246B1 publication Critical patent/EP1149246B1/fr
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

Links

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
    • F15B11/161—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load
    • F15B11/163—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors with sensing of servomotor demand or load for sharing the pump output equally amongst users or groups of users, e.g. using anti-saturation, pressure compensation
    • 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
    • F15B2211/31—Directional control characterised by the positions of the valve element
    • F15B2211/3105—Neutral or centre positions
    • F15B2211/3116—Neutral or centre positions the pump port being open in the centre position, e.g. so-called open centre
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/40—Flow control
    • F15B2211/405—Flow control characterised by the type of flow control means or valve
    • F15B2211/40515—Flow control characterised by the type of flow control means or valve with variable throttles or orifices
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/40—Flow control
    • F15B2211/405—Flow control characterised by the type of flow control means or valve
    • F15B2211/40553—Flow control characterised by the type of flow control means or valve with pressure compensating valves
    • F15B2211/40569—Flow control characterised by the type of flow control means or valve with pressure compensating valves the pressure compensating valve arranged downstream of the flow control means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/40—Flow control
    • F15B2211/42—Flow control characterised by the type of actuation
    • F15B2211/421—Flow control characterised by the type of actuation mechanically
    • F15B2211/423—Flow control characterised by the type of actuation mechanically manually, e.g. by using a lever or pedal
    • 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
    • F15B2211/57—Control of a differential 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
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/67—Methods for controlling pilot 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
    • F15B2211/70—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

Definitions

  • Control arrangement for at least two hydraulic consumers and pressure differential valve therefor
  • the invention relates to a control arrangement with which at least two hydraulic consumers are supplied with pressure medium and which has the features from the preamble of patent claim 1.
  • the invention also relates to a pressure differential valve which is used in particular in the control arrangement mentioned.
  • a hydraulic control arrangement is known for example from EP 0 566 449 A1. It is a hydraulic control arrangement based on the load-sensing principle, with i5 in which a variable pump is set depending on the highest load pressure of the hydraulic consumers operated so that the inlet pressure is a certain pressure difference above the highest load pressure.
  • the pressure medium flows to the two hydraulic consumers via two adjustable metering orifices, of which a first is arranged between a pump line coming from the variable displacement pump and a first hydraulic consumer and the second between the pump line and the second hydraulic consumer.
  • the pressure compensators connected downstream of the metering orifices ensure that with a sufficient quantity of pressure medium supplied, regardless of the load pressures of the hydraulic consumers, there is a certain pressure difference 5 across the metering orifices, so that the quantity of pressure medium flowing to a hydraulic consumer only depends on the opening cross section of the respective metering orifice. If a metering orifice is opened further, more pressure medium has to flow over it in order to generate the specific pressure difference. The variable pump is adjusted so that it delivers the required amount of pressure medium.
  • the pressure compensators downstream of the metering orifices are acted upon in the opening direction by the pressure after the respective metering orifice and in the closing direction by a control pressure present in a rear control chamber, which usually corresponds to the highest load pressure of all hydraulic consumers supplied by the same hydraulic pump. If, with simultaneous actuation of several hydraulic consumers, the metering orifices are opened so far that the quantity of pressure medium supplied by the hydraulic pump adjusted to the stop is smaller than the total quantity of pressure medium required, the quantity of pressure medium flowing to the individual hydraulic consumers becomes independent of the respective load pressure of the hydraulic Consumers reduced proportionately.
  • LUDV control Hydraulic consumers controlled in this way are called LUDV consumers for short.
  • a LUDV control is a special case of a load-sensing or load-sensing control (LS control) .
  • An advantage of a LS control with the pressure compensators connected upstream of the metering orifices compared to an LS control with the pressure compensators connected downstream of the metering orifices is, however, that the upstream pressure compensators by reducing the oversupply supplied for a short time by the variable displacement pump and an associated increase in the inlet pressure their opening cross-section do not allow the pressure difference across the metering orifices to increase, so that no more pressure medium flows through the metering orifices and the speed of the hydraulic consumers is not changed. The excess flows back to a tank via a pressure relief valve. In contrast, in the case of a control with the pressure compensators connected downstream of the metering orifices, the excess quantity is passed on to the hydraulic consumers.
  • the invention is based on the objective of designing a hydraulic control arrangement which has the features from the preamble of patent claim 1, in which, in particular, the pressure compensators are connected downstream of the metering orifices, in such a way that the inflow of excess quantities to the hydraulic consumers is prevented.
  • the desired aim is achieved according to the invention in that, in a generic hydraulic control arrangement according to the characterizing part of patent claim 1, the control pistons of the pressure compensators can be acted upon in the closing direction by a control pressure present in a rear control chamber which is derived from the supply pressure prevailing in the supply line with the aid of a valve device is and changes with the inlet pressure.
  • control pressure in the rear control chamber is derived from the inlet pressure and changes with it this. If the inlet pressure rises due to a delivery volume of the variable displacement pump that exceeds the demand, the control pressure also rises. Accordingly, the control pistons of the pressure compensators in
  • a constant pressure difference across an orifice plate means, however, with a constant opening cross section of the orifice plate, also means a constant amount of pressure medium flowing through the orifice plate.
  • the difference between the inlet pressure and the control pressure when the variable displacement pump has not yet been adjusted to the stop is not greater than between the inlet pressure and the highest load pressure. If the pressure difference were greater, the amount of pressure medium flowing to a hydraulic consumer would depend on whether the load pressure of this hydraulic consumer is higher or lower than the control pressure.
  • the control pressure is preferably slightly higher than the highest load pressure, so that on the one hand there are no unnecessary throttling losses on the pressure compensators, but on the other hand the pressure compensator assigned to the respective hydraulic consumer with the highest load pressure is still in the control range.
  • the pressure difference valve is preferably set to a fixed pressure difference and has a movable valve member which, in the sense of opening the fluidic connection between the inlet line and the control chamber on the pressure compensator, from the inlet pressure and in the sense of closing this connection from the control pressure and a spring is applied.
  • a particularly preferred embodiment also contains claim 5, according to which the rear control rooms of several pressure compensators are directly connected to one another, so that the same control pressure prevails in these control rooms. Only one valve device for deriving the control pressure from the inlet pressure is therefore necessary for these pressure compensators.
  • control arrangement has a load signal line, into which the highest load pressure of the hydraulic consumers actuated is given via selection valves, and a valve which opens a fluidic connection from the load signal line to the rear control chamber of at least one pressure compensator when the difference between the inlet pressure and the highest load pressure falls below a certain value.
  • a hydraulic consumer is to be supplied primarily with pressure medium compared to another hydraulic consumer in the event of undersaturation, this is advantageously done by an embodiment according to claim 8.
  • the rear control chamber on the pressure balance of the hydraulic consumer which is primarily to be supplied with pressure medium is then separated from the control rooms to the pressure balances of the other hydraulic consumers.
  • the control pressure in it is derived from the inlet pressure via a further valve device.
  • the priority valve which is used to maintain a desired pressure difference across the metering orifice arranged upstream of the pressure balance of the privileged hydraulic consumer and thus to maintain an adequate supply of pressure medium to the privileged hydraulic consumer when the hydraulic the corresponding delivery rate of the variable pump, the control pressure in the rear control chamber of the other hydraulic consumers can be raised above the control pressure in the event of saturation.
  • the priority valve according to claim 9 has a first connection connected to the supply line and a second connection connected to the rear control chambers of the pressure compensators assigned to the non-privileged hydraulic consumers and has a valve member which opens in the direction of the connection between the first connection and the second connection from in a line section downstream of the preferred hydraulic consumer
  • a control chamber of the priority valve can be connected upstream or downstream of the pressure compensator to the line section downstream of the metering orifice, since the priority valve is then in
  • FIG. 1 shows a circuit diagram of the exemplary embodiment of the control arrangement, which shows LUDV behavior in the event of undersaturation and which contains a privileged hydraulic consumer,
  • FIG. 1a shows an alternative to actuating the priority valve shown in FIG. 1,
  • FIG. 2 shows the circuit diagram of a variable displacement pump used in the exemplary embodiment
  • FIG. 3 shows a longitudinal section through the pressure differential valve used in the exemplary embodiment according to FIG.
  • a variable displacement pump 10 with a variable displacement 11 sucks pressure medium from a tank and releases it into a system of supply lines 13.
  • three hydraulic consumers 14, 15 and 16, all of which are designed as differential cylinders, are supplied with pressure medium via the feed lines.
  • each differential cylinder 14, 15 and 16 is assigned a metering orifice 17, 18 and 19 and a 4/3-way valve 20, 21 and 22, respectively.
  • a metering orifice and a directional control valve are each integrated in such a way that the direction of movement of the differential cylinder is predetermined by the actuation of a valve spool that is spring-centered in a central position in a certain direction, and the opening cross section is determined by the path that the valve spool is moved the metering orifice is determined.
  • a concrete constructive solution solution is referred to EP 0 566 449 A1 already mentioned.
  • the metering orifices 17, 18 and 19 are connected to the system of the feed lines 13.
  • a pressure compensator 23, 24 and 25 is arranged, the control piston of which is not shown in the opening direction from the pressure downstream of the respective metering orifice and in the closing direction of a control pressure prevailing in a rear control chamber 26 is applied.
  • the directional control valves 20, 21 and 22 each have two consumer connections 30 and 31 connected to pressure chambers of the respective differential cylinder, an inlet connection 32 which is connected to the outlet of the respective pressure compensator, and a return connection 33, from which a return line leads to the tank 12. In the middle position of a directional control valve, the two consumer connections are shut off and the inlet connection is connected to the tank connection.
  • control pistons of the pressure compensators 23, 24 and 25 are acted upon in the direction of closing in addition to a control pressure by a weak compression spring 34, which has a pressure of e.g. is equivalent to only 0.5 bar.
  • a weak compression spring 34 which has a pressure of e.g. is equivalent to only 0.5 bar.
  • the control rooms 26 and 27 of the two pressure compensators 23 and 24 are connected to one another via a channel 35, so that the same control pressure is always present in the two control rooms 26 and 27.
  • Change-over valves 36 are connected to the outputs of the pressure compensators 23, 24 and 25 or to the inlet connections 32 of the directional control valves, which are linked together in such a way that the highest load pressure of all actuated in a load signaling line 37, which leads to the adjustment 11 of the pump 10 Differential cylinder is present.
  • the load signaling line 37 leads to a control valve 39 with three connections, one of which is connected to an actuating cylinder 40 of the variable displacement pump 10.
  • Another connection of the control valve 39 is connected to a feed line 13 and the third connection to tank 12.
  • control piston of the control valve 39 is acted upon in the direction of a connection of the first connection to the second connection by the pressure in the feed line 13 and in the direction of a connection of the first connection to the third connection by the pressure in the load signaling line 37 and by a control spring 41.
  • Variable pumps and control valves according to the circuit diagram according to FIG. 2 are generally known and readily available on the market. It is therefore unnecessary to go into this in more detail. It should only be pointed out that the load-sensing or load-sensing pump control shown has the effect that a pressure is established in the feed line 13 which is a pressure difference equivalent to the force of the control spring 41 above the pressure in the load signaling line 37.
  • a pressure differential valve 45 is arranged between the system of the feed lines 13 and the channel 35 between the two control rooms 26 of the pressure compensators 23 and 24. This is connected to the inlet lines 13 with an inlet opening 46 and to the channel 35 with an outlet opening 47. Depending on the 0 position of a spool 48 of the pressure difference valve 45, which is not visible in FIG. 1 but is shown in FIG. 3, the inlet opening 46 and the outlet opening 47 are blocked off from one another or are fluidly connected to one another via a more or less large opening cross section.
  • the piston slide 48 is in the direction of reducing the opening cross section between the inlet opening and the outlet opening from the pressure prevailing in the channel 35 and in the control chambers 26 of the pressure compensators and from a compression spring 49 and in the direction of increasing the opening cross section from the inlet pressure prevailing in the feed lines 13 acted upon.
  • the effective areas on the spool for attacking the control pressure and the inlet pressure are equal in size, so that the pressure differential valve 45 ensures that the control pressure present in the channel 35 follows an increasing inlet pressure at a distance of a differential pressure equivalent to the force of the compression spring 49.
  • the pressure differential valve 45 is set so that the control pressure is 20 bar lower than the inlet pressure.
  • the channel 35 is connected to the tank 12 via a small flow regulator 50, so that the control pressure in the channel 35 can also follow a decreasing inlet pressure through the discharge of pressure medium via the small flow regulator 50.
  • a check valve 51 is connected between the load signaling line 37 and the channel 35 and opens from the load-reducing line 37 to the channel 35 when the pressure in the channel 35 becomes equal to the pressure in the load signaling channel 37.
  • the control pressure present in the control rooms 26 of the pressure compensators 23 and 24 cannot therefore fall below the highest load pressure present in the load reporting line 37.
  • a second pressure difference valve 52 which is identical to the pressure difference valve 45 and whose inlet opening 46 is also connected to a feed line 13.
  • the outlet opening 47 of the pressure differential valve 52 is connected to the control chamber 26 of the pressure compensator 25.
  • Both valves are set to the same pressure difference of, for example, 20 bar.
  • the control pressure in the control rooms 26 is 20 bar lower than the inlet pressure and, since this should be 25 bar higher than the highest load pressure, 5 bar higher than the highest load pressure.
  • the pressure compensators 23, 24 and 25 are therefore all in control, including those assigned to the consumer with the highest load pressure. Position.
  • the control chamber 26 of the pressure compensator 25 is connected to the tank 12 via a second small current regulator 50.
  • the differential cylinder 16 should, if the variable pump 10 brings maximum delivery rate and this does not meet the demand, be supplied with pressure medium primarily before the other two hydraulic cylinders 14 and 15.
  • a priority valve 55 is provided, which is designed as a proportional orifice with an inlet 56 and with an outlet 57. The latter is fluidly connected to the channel 35.
  • the inlet 56 is connected upstream of the metering orifice 19 to a feed line 13.
  • the movable valve member of the priority valve is acted upon in the direction of closing the connection between the inlet and the outlet by the pressure in the inlet, that is to say by the inlet pressure and in the direction of opening the connection by the pressure downstream of the metering orifice 19 and by the force of a control spring 58.
  • the control spring 58 is designed, for example, in such a way that a force equilibrium exists on the valve member of the priority valve if the pressure difference between the inlet pressure and the pressure downstream of the metering orifice 19 is 19 bar. This value is slightly less than the value of the pressure difference across the pressure difference valve 52 reduced by a pressure value of 0.5 bar equivalent to the force of the compression spring 34.
  • the priority valve 55 does not respond. If, by reducing the inlet pressure, the pressure difference across the metering orifice 19 drops below 19.5 bar, the pressure compensator 25 opens completely, so that the pressure downstream of the metering orifice 19 is equal to the load pressure of the primary hydraulic consumer 16. The load pressure of the consumer 16 is now present on the priority valve 55 on the spring side. He can open the priority valve 55 against the inlet pressure, whereby the pressure in the channel 35 and thus in the control rooms 26 of the pressure compensators 23 and 24 is raised above the highest load pressure.
  • the pressure compensators 23 and 24 are therefore adjusted in the closing direction until an increase in the pressure downstream of the metering orifices 17 and 18 an equilibrium of forces is reached on their control piston. However, the pressure difference across the metering orifices 17 and 18 is now reduced. The pressure medium flows flowing to the consumers 14 and 15 have become smaller. Ultimately, the priority valve 55 ensures by increasing the pressure in the 5 control rooms 26 of the pressure compensators 23 and 24 that by increasing the control pressure in the channel 35, the pressure difference across the metering orifices 17 and 18 and thus that flowing to the hydraulic consumers 14 and 15 Pressure medium flows are each reduced to such an extent that a quantity of pressure medium flows via the metering orifice 19, which produces a pressure difference which is approximately equal to the IO pressure difference in normal operation.
  • connection between the metering orifice 19 and the pressure compensator 25 can also be connected to the outlet of the pressure compensator 25, as shown in FIG. 1a.
  • the valve member of the priority valve 55 is then always acted upon by the load pressure of the priority hydraulic consumer 16 in the direction of opening the connection between the inlet 56 and the outlet 57.
  • the 0 priority valve can now be set to the same pressure difference that also prevails in normal operation via the metering orifice 19, since in normal operation the pressure difference between the load pressure of the priority hydraulic consumer 16 and the inlet pressure is higher than the pressure difference via the metering orifice 19 and therefore the priority valve 55 certainly does not respond. 5
  • the control pressure in the channel 35 becomes equal to the highest load pressure of the two hydraulic consumers 14 and 15 in the load signaling line 37 due to the drop in the inlet pressure the check valve 51 the highest load pressure also reported in the channel 35.
  • a further drop in the inlet pressure therefore no longer results in a further drop in the control pressure in the channel 35 and in the control spaces 26 of the pressure compensators 23 and 24.
  • This pressure which is slightly higher than the highest load pressure, is present downstream of both metering orifices 17 and 18. Upstream of both metering orifices 17 and 18 there is inlet pressure. Thus, the pressure difference across the metering orifice 17 is equal to the pressure difference across the metering orifice 18.
  • the pressure medium flows to the hydraulic consumers 14 and 15 are therefore reduced proportionately in the event of undersaturation regardless of whether the privileged consumer 16 is actuated. Consumers 14 and 15 are therefore LUDV consumers.
  • variable pump 10 If the pressure medium requirement of all simultaneously operated hydraulic consumers is covered by the variable pump 10, the pressure differential valves 45 and 52 together with the flow regulators 50 ensure that the control pressures in the control chambers 26 of the pressure compensators follow the inlet pressure at a fixed distance. If the variable displacement pump 10 now briefly produces an amount that exceeds what is required, for example because a wide-open metering orifice is completely closed, the inlet pressure rises sharply for a short time.
  • the control pressures follow this increase, so that the control pistons of the pressure compensators are subjected to an increased control pressure in the closing direction, move in the closing direction of the pressure compensators and thereby raise the pressure downstream of the metering orifices, so that the pressure difference across the metering orifices 17, 18 and 19 remains the same or increases only slightly. So the speed of a hydraulic consumer does not increase.
  • the excess flows through a pressure relief valve 60 to the tank.
  • the pressure difference valves 45 and 52 used in the control arrangement according to FIG. 1 are, as has already been pointed out, the same and, as can be seen from FIG. 3, are designed as built-in cartridges. They have a cartridge housing 70, through which a stepped valve bore 71 passes axially.
  • an adjusting screw 72 is screwed into the valve bore 71, through which the valve bore 71 is closed and which serves to support the control spring 49.
  • This control spring is located in the section of the valve bore 71 with the larger diameter into which the adjusting screw 72 is also screwed.
  • the control spring 49 is supported with its end facing away from the adjusting screw 72 on the piston slide 48, which is guided in the valve bore 71 in an axially movable manner.
  • the free space in the valve bore between the adjusting screw 72 and the piston slide 48 can be referred to as the spring space 75.
  • a star of radial bores 76 which form the outlet 47 of the pressure differential valve, opens into this.
  • the spool 48 is axially guided in the latter section of the valve bore 71 and has an annular groove 80 on the outside through which an annular space is created between it and the wall of the valve bore 71.
  • an axial blind bore 81 is made in the piston slide 48, which extends into the area of the annular groove 80 and is connected there via individual radial bores 82 to the annular groove 80.
  • Further radial bores 83 ensure an open fluidic connection between the bore 81 and the spring chamber 75 and thus also the outlet 47 then when the piston slide 48 abuts with its one end face against a stop of the adjusting screw 72.
  • the piston slide 48 has an outer shoulder 84 with which it can be pressed by the control spring 49 against the inner shoulder of the valve bore 71.
  • the annular groove 80 is located between the bore star of the radial bores 78 and the end face 79 of the cartridge housing 70. There is no opening cross section between the radial bores 78 and the annular groove 80.
  • the piston slide is on both sides of the annular groove 80 48 sealingly guided in the valve bore 71 so that the radial bores 78 are fluidly separated from the spring chamber 75 and the annular groove 80 from the space in front of the end face 79 of the valve housing. There is therefore no fluidic connection between the inlet 46 and the outlet 47 of the valve.
  • the spool 48 is acted upon by the inlet pressure from the end face 79 of the valve housing 70.
  • the compression spring 49 counteracts this and, on an area of the same size as the inlet pressure, the outlet pressure present in outlet 47.

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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)
  • Safety Valves (AREA)

Abstract

L'invention concerne un dispositif de commande qui sert à alimenter en fluide de pression au moins deux consommateurs hydrauliques et qui présente une pompe à cylindrée variable, régulée de façon à fournir la quantité requise de fluide de pression, le réglage de ladite pompe pouvant être modifié par un régulateur de pompe en fonction de la pression de charge maximale des consommateurs hydrauliques actionnés. Ce dispositif comprend également deux obturateurs de dosage réglables, dont le premier est placé entre une conduite d'alimentation sortant de la pompe à cylindrée variable et un premier consommateur hydraulique, et le deuxième est placé entre la conduite d'alimentation et un deuxième consommateur hydraulique. Ce dispositif comprend en outre deux balances de pression, dont la première est placée en aval du premier obturateur de dosage et la deuxième est placée en aval du deuxième obturateur de dosage, et dont le piston de régulation peut être sollicité par la pression, afin de provoquer une ouverture, en aval de l'obturateur de dosage respectif. Afin d'empêcher, dans un tel dispositif de commande, qu'une quantité excessive de fluide de pression de la pompe à cylindrée variable ne parvienne, pendant un bref laps de temps, aux consommateurs hydrauliques, les pistons de régulation des balances de pression peuvent être sollicités par une pression de commande appliquée dans une chambre de commande arrière, afin de provoquer une fermeture. Ladite pression de commande est dérivée de la pression d'alimentation qui règne dans la conduite d'alimentation, au moyen d'un dispositif à soupape, et varie avec la pression d'alimentation. L'invention concerne également une soupape à pression différentielle conçue de manière compacte, qui fait succéder, à une pression croissante à son entrée, une pression à sa sortie, avec une différence de pression fixe. Associée à une décharge réduite de la sortie, vers le réservoir, cette soupape à pression différentielle permet d'établir une différence de pression fixe entre la pression de sortie et la pression d'entrée. Une telle soupape est particulièrement appropriée pour être utilisée dans un dispositif de commande selon l'invention.
EP00906203A 1999-02-05 2000-01-15 Dispositif de commande pour au moins deux consommateurs hydrauliques, et soupape a pression differentielle correspondante Expired - Lifetime EP1149246B1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
DE19904616 1999-02-05
DE19904616A DE19904616A1 (de) 1999-02-05 1999-02-05 Steueranordnung für wenigstens zwei hydraulische Verbraucher und Druckdifferenzventil dafür
PCT/EP2000/000294 WO2000046513A1 (fr) 1999-02-05 2000-01-15 Dispositif de commande pour au moins deux consommateurs hydrauliques, et soupape a pression differentielle correspondante

Publications (2)

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EP1149246A1 true EP1149246A1 (fr) 2001-10-31
EP1149246B1 EP1149246B1 (fr) 2005-01-05

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EP00906203A Expired - Lifetime EP1149246B1 (fr) 1999-02-05 2000-01-15 Dispositif de commande pour au moins deux consommateurs hydrauliques, et soupape a pression differentielle correspondante

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Country Link
US (1) US6644025B1 (fr)
EP (1) EP1149246B1 (fr)
JP (1) JP4739529B2 (fr)
DE (2) DE19904616A1 (fr)
WO (1) WO2000046513A1 (fr)

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Also Published As

Publication number Publication date
DE19904616A1 (de) 2000-08-10
WO2000046513A1 (fr) 2000-08-10
DE50009158D1 (de) 2005-02-10
US6644025B1 (en) 2003-11-11
EP1149246B1 (fr) 2005-01-05
JP2002536599A (ja) 2002-10-29
JP4739529B2 (ja) 2011-08-03

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