EP0297401B1 - Dispositif de commande hydraulique - Google Patents

Dispositif de commande hydraulique Download PDF

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Publication number
EP0297401B1
EP0297401B1 EP88109858A EP88109858A EP0297401B1 EP 0297401 B1 EP0297401 B1 EP 0297401B1 EP 88109858 A EP88109858 A EP 88109858A EP 88109858 A EP88109858 A EP 88109858A EP 0297401 B1 EP0297401 B1 EP 0297401B1
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EP
European Patent Office
Prior art keywords
control
pressure
conduit
throttle
location
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
EP88109858A
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German (de)
English (en)
Other versions
EP0297401A2 (fr
EP0297401A3 (fr
Inventor
Rudolf Brunner
Heiner Feichtenbeiner
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Heilmeier and Weinlein Fabrik fuer Oel Hydraulik GmbH and Co KG
Original Assignee
Heilmeier and Weinlein Fabrik fuer Oel Hydraulik GmbH and Co KG
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Priority to AT88109858T priority Critical patent/ATE101899T1/de
Publication of EP0297401A2 publication Critical patent/EP0297401A2/fr
Publication of EP0297401A3 publication Critical patent/EP0297401A3/fr
Application granted granted Critical
Publication of EP0297401B1 publication Critical patent/EP0297401B1/fr
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B11/00—Servomotor systems without provision for follow-up action; Circuits therefor
    • F15B11/16—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
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B66—HOISTING; LIFTING; HAULING
    • B66F—HOISTING, LIFTING, HAULING OR PUSHING, NOT OTHERWISE PROVIDED FOR, e.g. DEVICES WHICH APPLY A LIFTING OR PUSHING FORCE DIRECTLY TO THE SURFACE OF A LOAD
    • B66F9/00—Devices for lifting or lowering bulky or heavy goods for loading or unloading purposes
    • B66F9/06—Devices for lifting or lowering bulky or heavy goods for loading or unloading purposes movable, with their loads, on wheels or the like, e.g. fork-lift trucks
    • B66F9/075—Constructional features or details
    • B66F9/20—Means for actuating or controlling masts, platforms, or forks
    • B66F9/22—Hydraulic devices or systems
    • 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
    • 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
    • F15B2211/30505—Non-return valves, i.e. 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
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
    • 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
    • F15B2211/3111—Neutral or centre positions the pump port being closed in the centre position, e.g. so-called closed 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
    • F15B2211/30—Directional control
    • F15B2211/32—Directional control characterised by the type of actuation
    • F15B2211/321—Directional control characterised by the type of actuation mechanically
    • F15B2211/324—Directional 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
    • F15B2211/30—Directional control
    • F15B2211/35—Directional control combined with flow control
    • F15B2211/351—Flow control by regulating means in feed line, i.e. meter-in control
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/60—Circuit components or control therefor
    • F15B2211/605—Load sensing circuits
    • F15B2211/6051—Load sensing circuits having valve means between output member and the load sensing circuit
    • F15B2211/6055—Load sensing circuits having valve means between output member and the load sensing circuit using pressure relief 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
    • F15B2211/615—Filtering 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
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/705—Output members, e.g. hydraulic motors or cylinders or control therefor characterised by the type of output members or actuators
    • F15B2211/7051—Linear output members
    • F15B2211/7052—Single-acting output members
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/705—Output members, e.g. hydraulic motors or cylinders or control therefor characterised by the type of output members or actuators
    • F15B2211/7051—Linear output members
    • F15B2211/7053—Double-acting output members
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/70—Output members, e.g. hydraulic motors or cylinders or control therefor
    • F15B2211/71—Multiple output members, e.g. multiple hydraulic motors or cylinders
    • F15B2211/7135—Combinations of output members of different types, e.g. single-acting cylinders with rotary motors

Definitions

  • the invention relates to a hydraulic control device of the type specified in the preamble of claim 1.
  • the pressure increase over the entire working range of the control element of the directional control valve is controlled to the same extent, so that there is essentially the same difference between the pressure in the pump line and the consumer pressure over the entire working range until the maximum delivery rate is reached.
  • the maximum pressure difference is only required to reach the maximum delivery rate in the control position end position of the control element of the directional control valve.
  • In control positions within the stroke of the Control element from the neutral position up to the vicinity of the control position end position is wasted energy due to the then unnecessarily high pressure difference, which can lead to heating and excessive mechanical wear of the pressure medium.
  • This control device also requires shuttle valves in the control line circuit in order to supply the pressure compensator with the highest load pressure from one of the directional valves provided.
  • each directional control valve is assigned its own inflow controller, which throttles depending on the load pressure as soon as the pressure compensator in the pump line sets too high a pressure due to the higher load pressure in another directional control valve. But this is an expensive additional effort.
  • a control pump is used as the pressure source provided that feeds the consumer via a pressure compensation valve upstream of the directional control valve.
  • a control line connected to the control pump and the pressure compensation valve branches off from the working line between the compensation valve and the directional control valve and contains a first throttle.
  • the control line is connected via a second throttle to the working line carrying the load pressure and additionally via a third throttle to the other working line leading to the tank.
  • the first and third throttles are adjustable.
  • the third throttle is throttled in opposite directions when the directional control valve is adjusted in the direction of the end position.
  • the signal pressure in the control line is made up of parts of the load pressure and parts of the compensation pressure prevailing upstream of the directional control valve.
  • the flow rate is not kept constant, but increased with increasing pressure and reduced with decreasing load pressure. This has a damping effect, so that the control device quickly reaches a steady state.
  • the inlet pressures of the second and third throttles By mixing the inlet pressures of the second and third throttles, no strict load dependency is possible. Even at low load pressures, the pressure in the working line upstream of the directional control valve is still a relatively small stroke of the directional control valve is significantly higher than the load pressure and this strong pressure increase is maintained over a relatively large stroke of the directional control valve, although a strong pressure increase only towards the end of the stroke of the directional control valve would be needed in its end position. This results in a waste of energy and an unnecessary mechanical load on the pressure medium.
  • the invention has for its object to provide a hydraulic control device of the type mentioned, which is characterized by improved energy efficiency and gentle treatment of the pressure medium.
  • the second throttling point in the control line circuit acts so that it first controls a certain low pressure increase via the pressure compensator, which ensures that the pump line pressure is above the consumer pressure, but only to the extent that this over an initial stroke with a lower delivery rate for the ensures proper load-independent movement of the consumer.
  • the pressure in the pump line raised in at least one stage to such an extent that the maximum delivery rate or the highest speed of the consumer can be reached without problems. This results in an improved use of energy and the advantage that the pressure medium is mechanically less stressed and is not heated because the pressure medium flows out via the pressure compensator with little flow resistance over the initial region of the stroke of the control element of the directional valve.
  • the second throttle point is divided into two parallel throttles, one of which is deactivated towards the end of the stroke of the control element of the directional control valve, so that there is then a higher flow resistance in the second control line, from which the stage in the pressure increase results.
  • the flow resistance is reduced because of the two throttles of the second throttle point that are then active, and the pressure increase is also lower.
  • the transition between the two stages of pressure increase is not noticeable on the consumer. It would also be conceivable to fork the second control line in more than two parallel lines and to provide a separate throttle in each line in order to achieve more than two stages in the pressure increase over the stroke of the control element.
  • the level of pressure increase can in this way be controlled as desired depending on the intended use of the hydraulic control device. Because the larger throttle of the second throttle point is separated toward the end of the stroke of the control element of the directional control valve, there is a significant increase in the pressure on the spring side of the pressure compensator. The procedure could also be reversed, so that the pressure increase stage is only lower.
  • the idea of the embodiment of claim 6 is also important, because it ensures that the inlet pressure of the first throttle point in the control line circuit on the side opposite the spring side of the pressure compensator becomes effective, that is to say that the pressure medium does not seek the lower-resistance route via the control branch channel can. If necessary, an additional influence on the course of the pressure increase can be exerted by appropriate coordination between the first throttle point and the throttle in the control branch channel or in the connecting channel.
  • a further, expedient embodiment of the subject matter of the invention in which several directional valves are connected in parallel to the pump line and the control line circuit, can be seen from claim 7.
  • the check valve not only prevents the load from dropping due to pressure medium possibly pushed back into the control line circuit, but also ensures that one which is actuated prematurely Directional control valve with a lower load pressure fed into the control line circuit when the directional control valve assigned to the check valve is subsequently switched on, which may control a higher load pressure, no step occurs in the pressure increase due to this higher load pressure, which would be dangerous for the other directional control valves or their consumers.
  • the check valve then uncouples the directional valve with the higher load pressure from the control line circuit and controls the pressure in the control line circuit with the lower, then priority load pressure.
  • This is an extremely important property of the control device, in particular in the case of forklift trucks or forklifts, because there the lifting cylinder usually works with the highest load pressure, while tilting cylinders or other auxiliary cylinders have to work with lower load pressures. There would be a dangerous interaction if the higher load pressure were not intercepted at the check valve.
  • the embodiment according to claim 8 is also expedient, because with this configuration of the check valve, it is assigned a double function by both forming the second throttle point and preventing the backward effect of the possibly disturbing high load pressure in the control line circuit. This is also a measure that is favorable in terms of production technology.
  • the embodiment according to claim 9 has proven to be particularly useful. From approximately 80% of the stroke of the control element of the directional control valve, a high pressure difference between the pump line pressure and the consumer line pressure is only required in practice. Previously, this high pressure differential only meant a waste of energy that is at the expense of the temperature of the pressure medium and increases its mechanical load unnecessarily.
  • a further, simplified embodiment finally emerges from claim 10.
  • This configuration results in a particularly expedient accommodation of the second throttle point, which also makes it possible to control the pressure increase for only one working direction of the consumer. It would also be conceivable to choose the cross-section of the tapping line so small that it functions as a second throttle point in the control line circuit.
  • a check valve 5 is provided in front of each directional valve 2, 3 and 3 '.
  • the directional control valves 2, 3, 3 ' are connected to a common return line 6 to a tank.
  • a pressure compensator 7 of conventional design is provided in the pump line 4, which contains a slide 8 which is infinitely adjustable between a shut-off position (FIG.
  • a control line circuit consists of a first control line 11, a second control line 12, a third control line 34 and a control line circuit part 30 connected to the first control line 11 at 29.
  • the first control line 11 branches off from the pump line 4 and leads to the directional control valve 2 and via it the other directional valves 3 and 3 'to the return.
  • An adjustable control element 17 contains one in each directional control valve Through channel 28, which connects the first control line 11 to the return line 6 in the neutral position.
  • the second control line 12 leads from the spring side of the pressure compensator 7 to a connection point 13 with the first control line 11 and from the connection point 13 to the directional control valve 2.
  • a first throttle point 14 is provided in the first control line 11 between the pump line 4 and the connection point 13.
  • a second throttle point 15 is provided in the second control line 12 between the crossing point 13 and the directional control valve 2.
  • the input pressure of the first throttle point 14 is transmitted by means of the third control line 34 to the side of the slide 8 of the pressure compensator 7 opposite the spring side.
  • the input pressure of the second throttle point 15 is effective via the second control line 12 on the spring side of the pressure compensator 7.
  • a check valve 16 is provided in the second control line 12 between the connection point 13 and the second throttle point 15 and is open in the direction of flow to the directional control valve 2.
  • the second control line 12 is forked behind the check valve 16 into two parallel branches 12a and 12b, each of which contains a throttle 15a and 15b forming part of the throttle point 15.
  • the two parallel branches 12a and 12b are connected to separate load pressure tapping connections 20, 21 in the directional control valve 2.
  • the throttle cross section of the first throttle point 14 is larger than the sum of the throttle cross sections of the throttles 15a and 15b.
  • the throttle 15b has a larger throttle cross section than the throttle 15a.
  • control member 17 is in conventionally adjustable from the neutral position in two control position end positions, with an intermediate position of the control member 17 is indicated in dashed lines in Fig. 1, in which this has performed less than, for example, 80% of the stroke in the direction of the first control position end position.
  • the directional control valve 2 is used to control a consumer, e.g. a single-acting cylinder 35, which in the present case can be the lifting cylinder of a forklift.
  • a consumer line 3c leads from the directional valve 2 to the cylinder 35.
  • a tap line 18 branches off to a load pressure tap connection 19 of the directional valve 2.
  • a connecting channel 22 is provided which bifurcates into two branches 24, 25 which are connected together to a channel part 23.
  • the fork load 24 or the connection to the connection 20 is shut off (indicated at 24a).
  • a connection channel 26 is provided with an adjustable throttle point 27, which leads the pressure medium from the pump line 4 into the consumer line 3c.
  • the channel 28 of the control member 17 is set to pass through the first control line 11 in the neutral position, and also in the second control line b, in which the pressure medium can flow out of the consumer line 3c directly into the return line 6.
  • An auxiliary control line 32 leads from line part 30 to a pilot-controlled pressure relief valve 33, with which the system pressure in the control line circuit is limited and that is connected to line 10 to return line 6.
  • a cylinder 36 Connected to the next directional control valve 3 as a consumer is a cylinder 36 which can be acted on on both sides and which has consumer lines 3a and 3b which can be acted upon alternately from the pump line 4.
  • throttles 31 are arranged in the directional control valve 3 in the consumer lines 3a and 3b, which limit the maximum delivery quantity, so that the cylinder 36 can only be moved at a limited speed.
  • a second throttle point 15' is also provided in the second control line 12 '.
  • the directional control valve 3 ' which corresponds to the directional control valve 3 except for the throttles 31, is of conventional design and is used to control a double-sided cylinder 37.
  • the second throttle point 15' is again provided in the second control line 12' to the directional control valve 3 '.
  • the throttle 15b and the throttle points 15 ' have, e.g. in this embodiment, the same throttle cross section.
  • the control device 1A works as follows: In the neutral position shown, all directional valves 2, 3 and 3 ', the pump line 4 is shut off. The pressure medium flowing into the first control line 11 reaches the return line 6 directly via the channels 28. The second control line 12 is thus relieved, so that the inlet pressure at the first throttle point 14 presses the slide 8 of the pressure compensator 7 into the open position via the third control line 34, as a result of which the pressure medium flows from the pump line 4 via the line 10 directly into the return line.
  • the pump essentially only needs to overcome the flow resistance generated by the weak spring 9.
  • the control element 17 of the directional control valve 2 is adjusted from the neutral position into an intermediate position (indicated by the broken line in FIG. 1) in the direction of the first control position end position, the passage of the first control line 11 open to the return line 6 is interrupted.
  • the connecting channel 22 connects the connections 19, 20 and 21.
  • the channel 26 of the control element 17 connects the pump line 4 to the consumer line 3c.
  • the load pressure in the consumer line 3c holds the check valve 16 in the blocking position.
  • a pressure builds up in the second control line 12, which moves the slide 8 of the pressure compensator in the direction of the shut-off position.
  • the input pressure of the second throttle point 15 is effective on the spring side of the pressure compensator 7, during the Input pressure of the first throttle point 14 acts on the opposite side of the slide 8.
  • the pressure compensator 7 regulates the speed of the cylinder 35 set with the control member 17 independently of the load pressure.
  • the input pressure of the second throttle point 15 results from the flow resistance of the two parallel throttles 15a and 15b, so that one of the pressure difference between the two input pressures and the spring 9 certain pressure increase results.
  • the fork load 24a of the connecting channel 22 is blocked.
  • the throttle 15b of the second throttle point 15 is thus effective.
  • the pressure medium in the second control line 12 only flows through the throttle 15a, so that the flow resistance increases and with it also the input pressure effective on the spring side of the pressure compensator 7.
  • the pressure in the pump line 4 is gradually increased in relation to the consumer pressure, so that the desired maximum delivery rate is finally reached up to the end position of the control member 17.
  • the throttle 27 in the control member 17 acts as a measuring throttle, while the pressure compensator 7 works as an adjusting throttle which controls the set speed of the consumer 35 independently of the load pressure.
  • the connecting channel 22 forms with the one throttle 15 a or with the two parallel throttles 15a and 15b a pressure stage D with which a step-like pressure increase is achieved.
  • the pressure compensator 7 operates depending on the input pressures of the first throttle point 14 and the second throttle point 15 'in the second control line 12' to the directional control valve 3.
  • the pressure increase remains approximately the same over the entire working range of the directional valve 3.
  • the maximum flow rate to cylinder 36 is limited in each working direction by throttles 31, e.g. to 30 l / min.
  • the lowest load pressure has priority over the higher load pressures. That is, leads, for example, the consumer line 3a of the directional control valve 3 the lowest load pressure, so the pressure compensator 7 works depending on the input pressure at the first throttle point 14 and the input pressure at the second throttle point 15 'of the directional control valve 3. Even if the load pressure at the check valve 16 would be higher or the load pressure at the directional control valve 3 ', this higher load pressure cannot affect the working of the pressure compensator 7, since it is reduced via the second throttle point 15 'at the directional control valve 3 relative to the level of the load pressure prevailing there.
  • the directional control valve 3 controls the tilt cylinder, the movement of which is to be maintained at a low speed, even when a large load is present on the lifting cylinder (directional control valve 2). Since the pressure prevailing in the pump line 4 then depends on the load pressure on the directional control valve 3, the cylinder 35 cannot move a high load and the speed of the cylinder 36 cannot be increased even via the high load pressure on the directional control valve 2, even if the user of the Forklift tries to outwit the hydraulic control device by operating the directional valve 2. Of course, the principle of gradually increasing the pressure, possibly even for every consumer direction, could be applied to all the directional valves provided.
  • the hydraulic control device 1B according to FIG. 2 differs from that of FIG. 1 by a modification of the pressure stage D for the input pressure of the second throttle point 15 or the spring side of the pressure compensator 7.
  • the other elements of the hydraulic control device 1B largely correspond to those described above that this will not be discussed any further.
  • the second control line 12 leads in Fig. 2 from the second throttle point 15, which is formed by a single throttle, the throttle cross section is smaller than that of the first throttle point 14, directly to the single load pressure tap connection 20 of the directional control valve 2 '. This is adjacent to the load pressure tap connection 19 of the tap line 18.
  • the control panel 17 ' is the Connection channel 22 'only from the branch 24 and the connection 23.
  • a continuous connection channel 40 is also provided in the control member 17', in the directional control valve 2 'on the side of the pump line 4, a control channel inlet connection 38 and opposite this a control channel outlet connection 39 are assigned so that the connections 38 and 39 are only connected at a predetermined stroke of the control member 17 'in the direction of the first control position end position, for example from 80% of the total stroke. In the stroke of the control member 17 'between the neutral position and this predetermined stroke (indicated by dashed lines), the connections 38 and 39 are separated.
  • a control branch channel 41 branches off from the pump line 4 upstream of the check valve 5 to the connection 38, and from the opposite connection 39 a connection channel 43 to an intersection 44 with the first control line 11.
  • the connection channel 43 could also be connected to another point in the control line circuit.
  • a throttle 42 is arranged in the control branch duct 41, the throttle cross section of which is equal to the throttle cross section of the first throttle point 14.
  • the tap line 18 is connected to the second control line 12 via the connecting channel 22'.
  • the input pressure of the first throttle point 14 acts on one side of the slide 8, while the input pressure of the second throttle point 15 acts on the spring side of the slide 8.
  • the pressure increase is effective as a function of the inlet pressures of the throttle points 14, 15, so that the pressure in the pump line 4 exceeds the pressure in the consumer line 3c by a predetermined amount.
  • the channel 40 connects the connections 38 and 39.
  • pressure medium is additionally guided into the control line circuit via the then open flow path 41, 42, 38, 40, 39, 43, 11.
  • the flow resistance at the first throttle point 15 increases, which on the spring side of the slide 8 of the pressure compensator 7 ensures that there is a greater restriction in the flow path from the pump line 4 to the line 10. This results in a growing pressure difference between the pressure in the pump line 4 and the consumer line 3c, which ensures that the maximum delivery rate in the consumer line 3c is reached.
  • the inlet pressure present at the first throttle point 14, which acts on the slide 8 against the spring 9, is not influenced by the pressure medium flowing into the control line circuit in the additional flow connection, because the throttle 42 has the same throttle cross section as the first throttle point 14.
  • connection 38 and 39 are separated again as soon as the control member 17' under e.g. 80% of the stroke travel. Then the lower pressure increase is activated until the neutral position is reached.
  • a preload check valve 45 is arranged in the second control line 12, which contains the second throttle point 15 on the one hand and the check valve 16 of the embodiment from FIG. 2 on the other hand united.
  • the connecting channel 22 'in the control element of the directional control valve corresponds to the connecting channel 22, which was explained with reference to FIG. 2, and which connects the tap line 18 both in an intermediate position of the control element and in the control position end position with the second control line 12.
  • a pressure stage is not shown in this embodiment.
  • Fig. 3A illustrates a detail variant in which a bias check valve 45 'is housed in the bleed line 18 from the consumer line 3c to the control member 17 of the directional valve. This is a structural simplification because the second control line 12 does not have to contain any organs responsible for the pressure increase.
  • a certain pressure drop can be set via the preload valve, for example 15 bar. If the bias valve 45, 45 'is designed so that the spring preload can be changed, the input pressure can be adapted to different conditions.
  • the effect of pressure stage D can be seen from the diagram in FIG. 4.
  • the stroke of the control member 17, 17 'of the directional control valve 2.2' is plotted in percent.
  • the delivery rate Q is indicated on the horizontal axis.
  • the lower solid curve P (3) indicates the pressure in the consumer line 3c, while the upper, solid curve P (4) shows the pressure in the pump line 4.
  • the dash-dotted curve T shows the course of the pressure in the pump line 4 of a conventional control device.
  • the hatched area F represents the energy saving due to the effect of pressure stage D.
  • the pressure P (3) in the consumer line initially increases when the delivery rate is initially small and then runs approximately linearly until the maximum delivery rate Q max is reached.
  • the pressure P (4) in the pump line 4 initially also increases when the delivery rate is small, in order to then run essentially constant with a higher value than the pressure in the consumer line 3c (first stage of the pressure increase).
  • the pressure stage becomes effective, whereupon the pressure P (4) in the pump line 4 and thus the pressure difference to the pressure P (3) in the consumer line 3c increases to a maximum value which before the maximum flow Q max is reached (second stage of pressure increase).

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Claims (10)

  1. Dispositif de commande hydraulique (1A, 1B, 1C), avec un distributeur (2, 2') au moins, monté en amont d'un consommateur (35), et dont l'organe de commande (17, 17') obture une conduite de consommateur (3c), au moins, dans une position neutre, et relie cette conduite dans deux positions de commande (a, b), alternativement, avec une tuyauterie de pompage (4) ou avec une conduite de retour (6), avec une source de pression raccordée à la tuyauterie de pompage (4), avec une balance de pression (7), raccordée à la tuyauterie de pompage (4) et munie d'un tiroir (8), chargé par un ressort (9) en direction de la position d'obturation, pour le retour direct du fluide sous pression, véhiculé par la source de fluide sous pression (P) et non requis par le consommateur (35), de la tuyauterie de pompage (4) dans la conduite de retour (6), avec un circuit de conduites pilotes en dérivation sur la tuyauterie de pompage (4) et présentant une première, une deuxième et une troisième conduite pilote (11, 12, 34, 30), la première conduite pilote (11) menant de la tuyauterie de pompage (4) à un raccordement de décharge, relié à la conduite de retour (6) dans la position neutre du distributeur (2, 2'), tandis que la deuxième conduite pilote (12) mène du côté ressort de la balance de pression (7) à un raccordement de prise de pression de charge (20, 21), au moins, du distributeur (2, 2'), et est reliée à la première conduite pilote (11), le raccordement de prise de pression de charge (20, 21) étant relié à la conduite de consommateur (3c) dans une position de commande (a), du moins, du distributeur (2, 2'), avec un premier point d'étranglement (14), disposé dans la première conduite pilote (11), en amont du point de jonction (13) avec la deuxième conduite pilote (12), dans le sens d'écoulement en direction du distributeur (2, 2'), et dont la pression d'entrée est transmise à un côté du tiroir (8) de la balance de pression (7), par l'intermédiaire de la troisième conduite pilote (34), et avec un deuxième point d'étranglement (15, 15'), disposé dans la deuxième conduite pilote (12) en aval du point de jonction (13), et dont la pression d'entrée peut être relevée par le déplacement de l'organe de commande (17, 17') à partir de la position neutre, et être transmise au côté ressort du tiroir (8) de la balance de pression (7), par l'intermédiaire de la deuxième conduite pilote (12), caractérisé en ce que la pression d'entrée du deuxième point d'étranglement (15) peut être relevée graduellement en deux niveaux au moins, au moyen de l'organe de commande ( 17, 17') du distributeur ( 2, 2'), en fonction de la course de cet organe en direction d'une position fin de course de commande et avant l'obtention de la position fin de course.
  2. Dispositif de commande hydraulique suivant la revendication 1, caractérisé en ce que la deuxième conduite pilote (12) est subdivisée en aval du point de jonction (13) en deux branches parallèles (12a, 12b), au moins, dont chacune comporte un étranglement (15a, 15b) en tant que partie du deuxième point d'étranglement (15), et mène à un raccordement de prise de pression do charge (20, 21), séparé, du distributeur (2), et en ce qu'un canal de jonction (22) de la prise (18), bifurquant en direction des deux raccordements de prise de pression de charge (20, 21), est prévu dans l'organe de commande (17) du distributeur (2), l'une des branches de fourche (24) de ce canal pouvant être obturée par l'organe de commande (17), avant que la première position fin de course de commande ne soit atteinte.
  3. Dispositif de commande hydraulique suivant les revendications 1 et 2, caractérisé en ce qu'un canal pilote en dérivation (41) mène de la tuyauterie de pompage (4) à un raccordement d'entrée du canal pilote (38) du distributeur (2'), en ce qu'une conduite de jonction (43) est prévue d'un raccordement de sortie du canal pilote (39) du distributeur (2') en direction de la première conduite pilote (11) ou du circuit de conduites pilotes, et en ce qu'un passage (40) est prévu dans l'organe de commande (17') du distributeur (2'), ce passage reliant les raccordements du canal pilote (38, 39), lors de la course de l'organe de commande (17'), en avance de l'obtention de la position fin de course de commande.
  4. Dispositif de commande hydraulique suivant la revendication 2, caractérisé en ce que les deux étranglements (15a, 15b), prévus dans les branches parallèles (12a, 12b), ont une dimension différente, en ce que la somme des sections des deux étranglements (15a, 15b) est inférieure à la section du premier point d'étranglement (14), et en ce que la branche de fourche (24) du canal de jonction (22), menant en direction du plus gros (15b) des deux étranglements (15a, 15b), peut être obturée par l'organe de commande (17).
  5. Dispositif de commande hydraulique suivant la revendication 4, caractérisé en ce que le premier point d'étranglement (14) a un diamètre de 0,6 mm, les deux étranglements (15a, 15b) du deuxième point d'étranglement (15) ayant un diamètre de 0,5 et 0,3 mm.
  6. Dispositif de commande hydraulique suivant la revendication 3, caractérisé en ce qu'un étranglement (42), dont la section est égale à celle du premier étranglement (14), est disposé dans le canal pilote en dérivation (41) ou dans la conduite de jonction (43).
  7. Dispositif de commande hydraulique suivant l'une quelconque au moins des revendications 1 à 6, plusieurs distributeurs (2, 2', 3, 4) étant raccordés en parallèle à la tuyauterie de pompage (4) et au circuit de conduites pilotes (11, 12, 30), caractérisé en ce qu'un clapet antiretour (16), se fermant à l'inverse du sens d'écoulement en direction du distributeur (2, 2'), est disposé dans la deuxième conduite pilote (12), entre le point de jonction (13) et le deuxième point d'étranglement (15), pour le distributeur (2, 2'), du moins, prévu pour le consommateur (35) à alimenter par la pression maximale prévisible.
  8. Dispositif de commande hydraulique suivant l'une des revendications 3 et 7, caractérisé en ce que le deuxième point d'étranglement (15) est formé par un clapet antiretour à précontrainte, chargé par ressort, ajustable de préférence, qui remplace le clapet antiretour (16).
  9. Dispositif de commande hydraulique suivant l'une quelconque au moins des revendications 1 à 8, caractérisé en ce que la pression d'entrée du deuxième point d'étranglement (15, 15') peut être relevée au second niveau, à partir de 80% environ de la course de l'organe de commande (17, 17') en direction de la première position fin de course de commande.
  10. Dispositif de commande hydraulique suivant l'une quelconque au moins des revendications 1 à 9, caractérisé en ce que le deuxième point d'étranglement (15) est disposé dans une conduite de prise de pression de charge (18), entre la conduite de consommateur (3c) et l'organe de commande (17) du distributeur, en forme d'un clapet antiretour à précontrainte (45') chargé par ressort, de préférence.
EP88109858A 1987-07-03 1988-06-21 Dispositif de commande hydraulique Expired - Lifetime EP0297401B1 (fr)

Priority Applications (1)

Application Number Priority Date Filing Date Title
AT88109858T ATE101899T1 (de) 1987-07-03 1988-06-21 Hydraulische steuervorrichtung.

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE3722083A DE3722083C1 (de) 1987-07-03 1987-07-03 Hydraulische Steuervorrichtung
DE3722083 1987-07-03

Publications (3)

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EP0297401A2 EP0297401A2 (fr) 1989-01-04
EP0297401A3 EP0297401A3 (fr) 1991-03-13
EP0297401B1 true EP0297401B1 (fr) 1994-02-23

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EP88109858A Expired - Lifetime EP0297401B1 (fr) 1987-07-03 1988-06-21 Dispositif de commande hydraulique

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Country Link
US (1) US4879945A (fr)
EP (1) EP0297401B1 (fr)
JP (1) JPS6426002A (fr)
AT (1) ATE101899T1 (fr)
DE (2) DE3722083C1 (fr)

Families Citing this family (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE3807583C1 (fr) * 1988-03-08 1989-03-09 Heilmeier & Weinlein Fabrik Fuer Oel-Hydraulik Gmbh & Co Kg, 8000 Muenchen, De
US4986071A (en) * 1989-06-05 1991-01-22 Komatsu Dresser Company Fast response load sense control system
US5081839A (en) * 1990-01-29 1992-01-21 Caterpillar Inc. Pressure compensated hydraulic system
SE469485B (sv) * 1991-11-29 1993-07-12 Jan Lindholm Foerfarande och floedesregleringsventilaggregat foer floedesbalansering
GB2501486A (en) * 2012-04-24 2013-10-30 Jc Bamford Excavators Ltd Work machine having a hydraulic system comprising variable orifice ratios
CN103754798B (zh) * 2014-02-19 2017-06-20 上海梯佑叉车有限公司 一种半电动自升高装卸车

Citations (1)

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Publication number Priority date Publication date Assignee Title
US3971216A (en) * 1974-06-19 1976-07-27 The Scott & Fetzer Company Load responsive system with synthetic signal

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US3411295A (en) * 1967-05-31 1968-11-19 Gen Signal Corp Hydraulic supply systems
US3815477A (en) * 1973-02-06 1974-06-11 Cross Mfg Inc Control valve instrumentality
DE2457451A1 (de) * 1974-12-05 1976-06-10 Bosch Gmbh Robert Hydraulische steuereinrichtung
DE2601484C2 (de) * 1976-01-16 1984-07-26 Robert Bosch Gmbh, 7000 Stuttgart Steuereinrichtung für mindestens zwei von einer Druckmittelquelle gespeiste hydraulische Verbraucher
DE2804045A1 (de) * 1978-01-31 1979-08-09 Bosch Gmbh Robert Steuervorrichtung fuer einen hydraulisch betriebenen verbraucher
EP0167818B1 (fr) * 1984-07-10 1988-06-01 Robert Bosch Gmbh Dispositif de commande hydraulique
DE3425303A1 (de) * 1984-07-10 1986-01-23 Robert Bosch Gmbh, 7000 Stuttgart Hydraulische steuereinrichtung
DK154169C (da) * 1984-10-03 1989-03-20 Danfoss As Styreindretning for en hydraulisk drevet forbruger
DE3532816A1 (de) * 1985-09-13 1987-03-26 Rexroth Mannesmann Gmbh Steueranordnung fuer mindestens zwei von mindestens einer pumpe gespeiste hydraulische verbraucher
DE3611244A1 (de) * 1986-04-04 1987-10-08 Rexroth Mannesmann Gmbh Stromregelventil

Patent Citations (1)

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Publication number Priority date Publication date Assignee Title
US3971216A (en) * 1974-06-19 1976-07-27 The Scott & Fetzer Company Load responsive system with synthetic signal

Also Published As

Publication number Publication date
DE3887932D1 (de) 1994-03-31
DE3722083C1 (de) 1988-09-15
EP0297401A2 (fr) 1989-01-04
EP0297401A3 (fr) 1991-03-13
JPS6426002A (en) 1989-01-27
US4879945A (en) 1989-11-14
ATE101899T1 (de) 1994-03-15

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