EP2241764B1 - Soupape à siège dotée d'une fonction de soupape de dérivation et de compensateur de pression - Google Patents

Soupape à siège dotée d'une fonction de soupape de dérivation et de compensateur de pression Download PDF

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Publication number
EP2241764B1
EP2241764B1 EP09005478A EP09005478A EP2241764B1 EP 2241764 B1 EP2241764 B1 EP 2241764B1 EP 09005478 A EP09005478 A EP 09005478A EP 09005478 A EP09005478 A EP 09005478A EP 2241764 B1 EP2241764 B1 EP 2241764B1
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EP
European Patent Office
Prior art keywords
pressure
valve
control piston
line
seating valve
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.)
Not-in-force
Application number
EP09005478A
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German (de)
English (en)
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EP2241764A1 (fr
Inventor
Harald Klemens
Josef Poldinger
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.)
Hawe Hydraulik SE
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Hawe Hydraulik SE
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Publication date
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Priority to EP09005478A priority Critical patent/EP2241764B1/fr
Priority to AT09005478T priority patent/ATE522728T1/de
Publication of EP2241764A1 publication Critical patent/EP2241764A1/fr
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Publication of EP2241764B1 publication Critical patent/EP2241764B1/fr
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/024—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
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0401—Valve members; Fluid interconnections therefor
    • F15B13/0405—Valve members; Fluid interconnections therefor for seat valves, i.e. poppet 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
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0416—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor with means or adapted for load sensing
    • F15B13/0417—Load sensing elements; Internal fluid connections therefor; Anti-saturation or pressure-compensation valves

Definitions

  • the invention relates to a seat valve according to the preamble of patent claim 1.
  • Hydraulic control systems with pressure and volume flow adaptation to the current requirements of one or more consumers are realized in so-called LS systems.
  • Such LS systems can be realized, for example, with constant pump and 3-way pressure compensator.
  • the highest pressure occurring in the system in the lines to the actuators is compared to an inlet pressure compensator, such as the 3-way pressure compensator, and compared with the system pressure that the pump currently supplies.
  • the system pressure and volume flow is then regulated according to demand.
  • the unused volume flow with system pressure is discharged via the pressure compensator to the tank. If no hydraulic consumers are switched on, there is no pressure requirement and the entire pressure medium flow is diverted from the pressure compensator into the reservoir.
  • the 3-way pressure compensator acts as a circulation valve.
  • 3-way pressure compensators for example, are realized with a 2/2-way proportional valve with a hydraulic control input, which is therefore sometimes referred to as a 3/2-way valve.
  • Fig. 1 shows a hydraulic equivalent circuit diagram of such usable for this purpose 3/2-way proportional seat valve.
  • Fig. 1 shows a symbolic representation of the three-way seat valve V according to the prior art.
  • the three-way seat valve V has an input for connection to a pressure line P and an output for connection to a return line R.
  • a control input LS, on the z. B. the load pressure of a consumer is pressing the slide of the valve V in position b, so that the connection between the pressure line P and the return line R is interrupted.
  • a spring 17 supports the force effect of the pressure in the control line LS.
  • the pressure of the pressure line P is directed to the control piston of the valve V that he Compressive force from the control input LS and the spring force of the spring 17 counteracts. In FIG. 1 this is done symbolically by the line 19.
  • the line 19 may be a valve internal or external conduction path or may be omitted due to the design.
  • the spring force of the spring 17 is selected so that, when the control input LS is unpressurized, the control piston is pushed into position a due to the pressure in the pressure line P, so that a connection is established between the pressure line P and the return line R.
  • hydraulic actuators load control pressure equal to zero
  • the overpressure in the pressure lines to the consumers will be reduced to near zero when no hydraulic loads are applied (load control pressure equal to zero).
  • load control pressure load control pressure equal to zero
  • the opening cross section of the passage between the pressure line P and the return line R can be varied proportionally to the pressure on the control line LS, so that a proportional valve function can be realized.
  • the valve described above for example, in the European patent application EP-A-0965763 disclosed.
  • the EP-A-0965763 shows a hydraulically actuated 2/2-idle directional seat valve V, which is used as a circulation valve, and having the features of the introductory part of claim 1.
  • Fig. 2 a sectional view of this valve V of EP-A-0965763
  • Fig. 2 shows a pump line section P and a return line section R and a connection 12 between these line sections.
  • the directional seat valve V includes a valve seat 14 and at the return side a closing member 15, for example a ball.
  • the closing member 15 is acted upon in the return line section R by a piston 16, which is sealed sealed leakage in a chamber 18.
  • the piston 16 is loaded in the closing direction of the closing member 15 by a weak spring 17 and by the load pressure of one of the control lines 13.
  • a disadvantage of this type of valve is that, when the control piston of the valve V is in the orbital position (position a), only the back pressure of the pump, which arises when the pressure medium is pumped directly into the reservoir with the least possible line resistance in the circulation position, must be overcome before the spool can enter a proportional control range. Ie., If consumers are switched on from the circulation position with low pressure requirement, the pressure in the load control circuit is not enough to move the valve out of the circulation position.
  • hydraulic component component manufacturers are provided with ready-to-use valve combinations enabling two-stage recirculation operation, high-speed recirculation mode pressure change, and lower sensitivity to recirculation mode pressure changes.
  • HAWE Hydraulik SE offers a PSL.U connection block for connecting a pump to a large number of hydraulic consumers, which combines several valve functions.
  • the connection block is used for load-independent, stepless control of the speed of movement of hydraulic consumers.
  • Several consumers can be driven simultaneously and independently.
  • the field of application of this valve type is mainly in the field of mobile hydraulics (for example crane controls, etc.).
  • the automatic reduction of the pump pressure depending on the consumer demand and an automatic activation of the system for reducing the Umlaufwiederstandes and the function shutdown with a circulation valve guarantee a good adaptation to the respective control tasks. In this case, the control operation and the circulation operation of separate valves are carried out in a valve block.
  • FIG. 3 a hydraulic diagram of the PSL.U connection block from HAWE Hydraulik SE is shown.
  • Fig. 3 shows as essential components a 3-way pressure compensator 1 and a self-switching circulation valve 2.
  • the pressure compensator 1 operates in an internal control loop as a differential pressure regulator whose movable control edge changes so that the pressure difference between the pump pressure, which is applied to the control input 1-1 , and load pressure (load-sensing pressure) applied to the control input 1-2 is always constant regardless of the load pressure.
  • the control input 1-2 is a damping element 4 for improving the Control characteristic upstream.
  • a pressure relief valve 3 limits the maximum pressure at the control input 1-2.
  • the pressure compensator 1 lowers the pressure on the pressure line P depending on the load pressure on the load pressure line LS. If no consumer is active, the load pressure line LS is depressurized and the slide in the pressure compensator 1 is pushed against the spring force of the spring 1-3 in the position b, so that hydraulic fluid can flow from the pressure line directly via the pressure compensator 1 in the return and the pressure on the pressure line is thereby reduced. If the pressure on the load pressure line increases, the slide is increasingly pushed in the direction of position a, so that the pressure drop on the pressure line P is lower.
  • the hydraulically actuated self-switching circulating valve 2 is held in the rest position by a spring 2-3, supported by the pressure in the load-sensing line LS at the control input 2-2, in position a (closed state). If no consumers are active and the load-sensing line LS is depressurized, the valve 2 switches into circulation position. If hydraulic consumers are switched on again, the valve 2 switches back into blocking position, so that the proportional valve 1 can take over the control again.
  • a desired high sensitivity to pressure changes at the control input 2-2 can be achieved.
  • valve blocks A disadvantage of such valve blocks, however, is the complicated structure and the high production costs, since two independently acting control piston must be used.
  • a poppet valve is provided with a control piston for interrupting a connection between an input and an output, wherein the input is provided for connection to a pressure line and the output for connection to a line with lower pressure than the pressure in the pressure line, wherein a first pressure of the pressure line on a first loading surface of the control piston and wherein a second pressure in a control line acts on a second loading surface of the control piston in the opposite direction to the force acting on the first loading surface, and wherein the first and second loading surface are of different sizes.
  • the seat valve acts as a differential pressure regulator, as is known in the prior art.
  • a first channel is provided in the seat valve, which directs the first pressure of the pressure line to a third loading surface, which is opposite to the first loading surface.
  • a second channel is provided in the poppet valve to direct the first pressure from the pressure line to the second impingement surface in an open state of the poppet valve.
  • valve in the open state has a higher sensitivity to pressure changes at the control input than in the closed state.
  • the valve can be adjusted so that in orbit almost no back pressure has to be overcome to operate the valve. Complicated valve combinations as described in the prior art are not necessary.
  • the first impingement area is larger than the second and third impingement areas to achieve appropriate differential pressure control characteristics.
  • first channel and the second channel are formed in the control piston.
  • the channels are necessary to bring pressure medium to different loading surfaces.
  • the arrangement of the channels in the control piston has manufacturing advantages.
  • closing edges of the control piston are designed so that the passage cross section between input and output in dependence on a pressure difference between the first pressure of the pressure line and the second pressure in the control line can change continuously in accordance with a predetermined characteristic.
  • This control characteristics can be set individually depending on the design of the control edges on the control piston
  • the first channel is a first bore through the control piston parallel to a direction of movement of the control piston
  • the second passage is a second bore oblique or perpendicular to the direction of movement of the control piston which breaks through a side wall of the control piston to correspond to the pressure in the pressure line to bring to the correct pressure areas.
  • control piston is biased by a spring in the closing direction.
  • the switching characteristic of the valve can be influenced by the spring when it switches from the circulation mode to the regular mode. Depending on the conditions of the three pressurization surfaces, the switching and control characteristics can be adjusted via the spring.
  • the spring force of the spring is designed to compensate for frictional forces. If the control piston is designed so that the first surface is as large as the sum of the second and third surfaces only frictional forces must be overcome to change the valve position from the open state to the control position. To increase the switching sensitivity, the frictional forces are compensated by the spring.
  • first, second, and third apply surfaces are associated with a respective first, second, and third pressure chambers within the seat valve, wherein the first and third pressure chambers are fluidly interconnected by the first channel.
  • the valve housing In order for the pressurization surfaces to exert effective forces on the control piston, the valve housing must be designed so that the spatially separated pressure surfaces and spatially separated pressure chambers are assigned.
  • the second pressure chamber in the poppet valve is arranged so that the second channel moves towards the second pressure chamber when the passage between the inlet and outlet opens.
  • the relative position of the second channel in the control piston is set to obtain the desired 2-stage characteristic of the valve.
  • the second channel in the control piston is arranged so that it establishes a connection between the second pressure chamber and the first or third pressure chamber, when the control piston reaches at least 70% of the maximum stroke in the opening direction, wherein at maximum stroke, the passage between input and output is fully open.
  • the absolute position of the second channel is set to achieve the desired two-stage operation of the valve.
  • other lift heights e.g. 80% or 90%, be required, from which the connection is made.
  • Fig. 4 shows a symbolic representation of the three-way seat valve 100 according to the present invention.
  • Fig. 4 symbolizes a two-stage three-way valve 100 that combines two three-way valves 100a and 100b.
  • the valve member 100a corresponds schematically to the three-way valve of the prior art of Fig. 1 .
  • the valve portion 100a of FIG Fig. 4 an input for connection to a pressure line P and an output for connection to a return line R.
  • a control input LS, on the z. B. the load pressure of a consumer is pressing the slide of the valve 100 in position b, so that the connection between the pressure line P and the return line R is interrupted.
  • a spring 110 supports the force effect of the pressure in the control line LS.
  • the pressure of the pressure line P is directed to the control piston of the valve 100 so that it counteracts the pressure force from the control input LS and the spring force of the spring 110.
  • the spring force of the spring 110 is selected such that, when the control input LS is unpressurized, the control piston is pushed into position a due to the pressure in the pressure line P, so that a connection is established between the pressure line P and the return line R.
  • hydraulic actuators will not pump hydraulic loads (load control pressure equal to zero) directly back into the reservoir and the pressure in the pressure lines to the consumers will be reduced to near zero. As a result, the energy consumption can be reduced by circulating pressure losses.
  • the opening cross section of the passage between the pressure line P and the return line R can be varied proportionally to the pressure on the control line LS, so that a proportional valve function can be realized. This is represented symbolically by the two lines on the side of the valve section 100a.
  • the proportional valve reduces the cross section of the passage between the pressure line P and the return line R, ie the control piston shifts more in Direction position b, so that a larger volume flow for the pressure lines to the consumers is available.
  • the three-way valve 100 in Fig. 4 is further extended to the valve member 100b, which is symbolically represented by an extension of the spool with a smaller diameter.
  • the expansion of the spool about a piston part with a smaller diameter results in two surfaces F2 and F3, which can be pressurized.
  • the surface F1 opposite the surfaces F1 and F2 can be uniformly pressurized (here with the pressure from the pressure line P) exerting a force on the spool opposite to the force of the pressures on the surfaces F2 and F3.
  • the surface F2 is acted upon by the load pressure from the control input LS.
  • the surface F3 as well as the surface F1 is acted upon by the pressure from the pressure line P.
  • the valve 100 moves in position a and the valve 100 is in a circulating position, in which the pressure lines are relieved to the consumers.
  • the connection 140 in position a of the dynamic pressure in the pressure line P is directed to the surface F2, whereby the initial resistance is greatly reduced to bring the valve 100 from the orbital position a in the proportional directional range. If the slider was moved out of position a, z. B. by supplying pressure on the control line LS, the valve moves gradually in position b and the connection in the pressure line P and the pressurizing surface F2 is interrupted, so that only the load pressure at the control input LS of the surface F2 is acted upon.
  • the valve 100 has two sensitivity levels. As long as the valve is in the vicinity of the position a and the pressure of the pressure line P is directed to the pressurizing surface F2, the valve 100 is sensitive to lower pressure changes of the differential pressure between the pressure in the control line LS and the pressure line P. Once the connection 140 is interrupted, since the control piston has moved in the direction B, the control sensitivity is reduced because the pressure of the pressure line P is no longer on the pressurization surface F2, which acts as a control surface lies.
  • the spring force of the spring force 110 serves to further reduce the initial resistance for moving the valve from the circulation position A into the control range B. In particular, the spring 110 can be used to compensate for frictional forces. However, the valve remains functional even without spring 110.
  • Fig. 5 shows in a particular embodiment, a realization of the three-way valve 100, which is shown schematically in FIG Fig. 4 was shown.
  • Fig. 5 shows by reference numeral 150 a valve seat with a cylindrical bore having a first diameter, in which a control piston 120 is arranged to match the first diameter movable in the longitudinal direction to the bore along an axis X (symmetry longitudinal axis of the bore).
  • the valve seat 150 further has an input bore E, which is formed axially to the axis X at the opposite end of the valve seat 150 to the cylindrical bore, and which has a second diameter which is smaller than the first diameter of the cylindrical bore.
  • the input bore E is designed so that it can be connected to a pressure line.
  • the valve seat 105 has a further bore A, which is designed radially to the cylindrical bore (to the axis X) and which forms a passage through the lateral valve seat wall to the cylindrical bore.
  • the axial bore E of the second diameter forms a passage through a bottom surface of the valve seat 150 and strikes the cylindrical bore.
  • the seating surfaces between the control piston and the bottom of the valve seat 150 serve as a seal between the bore E and the bore A.
  • Specially shaped control edges 200 at the bottom of the control piston 120 are formed so that the cross-sectional area of the passage between the bore E and the bore A. continuously changes when the control piston 110 moves relative to the valve seat 150 in the cylindrical bore.
  • the control piston 120 has an axial to the axis X bore, which in the in Fig. 5 is shown open upwards, and in which a spring 110 is inserted.
  • At the bottom of the control piston 120 is also an axially to the axis X, which is also a piston axis of symmetry, located bore 130, which is a Input chamber F1A at the input bore E with a valve chamber F3A, in which the spring 110 is connected to each other.
  • a lid 160 is screwed on the valve seat 150 such that the control piston 120 is movable in the axial direction to the axis X by at least a maximum movement stroke H0 for opening and closing the connection between the input bore E and the output bore A.
  • the input bore E defines a surface F1, via which the control piston 120 can be pressurized.
  • the control piston 120 has a step along its longitudinal direction, above which the diameter of the control piston 120 is reduced.
  • the area at the stage of the control piston 120 defines a second pressure area F2.
  • the area on the face of the portion of the smaller diameter control piston 120 defines a third pressurizing area F3.
  • the valve seat 150 is formed so as to form a chamber F2A together with the step of the control piston 120.
  • a control port LS for connection to a load pressure control circuit (load sensing) is provided.
  • In the control piston 120 is located below the step defining the surface F2, an axial bore 140 which breaks through the side wall of the control piston.
  • the control piston 120 moves in the in Fig. 5 As shown, the axial bore 140 is at about 84% of the maximum lift height H0 in communication with the chamber F2A, so that the pressure at the input bore E can act on the second pressurization surface F2.
  • the valve 100 is close to fully open and the valve will then have an increased sensitivity to load pressure changes, as already associated with FIG FIG. 4 has been described.
  • Various seals 101 seal the different parts of the valve 100 to each other and to adjacent housings and conduit components.
  • the valve 100 is shown so that it z. B. can be used as a plug valve.
  • Fig. 5 also shows a spring plate 170 which is fixed to the screw 160, and whereby the spring 110 is fixed to the valve housing.
  • a pressure-tight seal 180 (seal-lock) of the valve housing is provided.
  • Fig. 5 shows a concrete execution, with which all functions, in connection with Fig. 4 described, can be realized.
  • the valve of Fig. 5 a 2-stage characteristic.
  • Same reference numerals in the FIGS. 4 and 5 means that these elements perform the same function.
  • the explanatory notes to the Fig. 4 and 5 are therefore to be understood as complementary and can be combined freely.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Safety Valves (AREA)
  • Fluid-Driven Valves (AREA)

Claims (12)

  1. Soupape à siège (100) comprenant un piston de commande (120) pour interrompre une liaison entre une entrée (E) et une sortie (A), soupape dans laquelle l'entrée (E) est prévue pour le raccordement à une conduite sous pression (P) et la sortie (A) est prévue pour le raccordement à une conduite sous pression plus basse que la pression dans la conduite (P), une première pression de la conduite sous pression (P) agit sur une première surface de sollicitation (F1) du piston de commande (120) et une seconde pression dans une conduite pilote (LS) agit sur une deuxième surface de sollicitation (F2) du piston de commande (120) dans la direction opposée à l'effet dynamique sur la première surface de sollicitation (F1), et la première et la deuxième surfaces de sollicitation (F1, F2) sont de dimension différente,
    caractérisée en ce que
    un premier canal (130) est prévu dans la soupape à siège (100), lequel dirige la première pression de la conduite sous pression (P) sur une troisième surface de sollicitation (F3), qui est en vis-à-vis de la première surface de sollicitation (F1), et que
    un second canal (140) est prévu dans la soupape à siège (100), lequel dirige, dans un état d'ouverture de la soupape à siège (100), la première pression de la conduite sous pression (P) sur la deuxième surface de sollicitation (F2).
  2. Soupape à siège (100) suivant la revendication 1, dans laquelle la première surface de sollicitation (F1) est plus grande que la deuxième surface de sollicitation (F2).
  3. Soupape à siège (100) suivant l'une des revendications 1 et 2, dans laquelle la première surface de sollicitation (F1) est plus grande que la troisième surface de sollicitation (F3).
  4. Soupape à siège (100) suivant l'une au moins des revendications 1 à 3, dans laquelle le premier canal (130) et le second canal (140) sont réalisés dans le piston de commande (120).
  5. Soupape à siège (100) suivant l'une au moins des revendications 1 à 4, dans laquelle des arêtes de fermeture (200) du piston de commande (120) sont réalisées de sorte que la section transversale de passage entre l'entrée (E) et la sortie (A) peut être modifiée en continu, suivant une caractéristique prédéfinie, en fonction d'une différence de pression entre la première pression de la conduite sous pression (P) et la seconde pression dans la conduite pilote (LS).
  6. Soupape à siège (100) suivant l'une au moins des revendications 4 et 5, dans laquelle le premier canal (130) est un premier alésage au travers du piston de commande (120) parallèlement à une direction de déplacement du piston de commande (120).
  7. Soupape à siège (100) suivant l'une au moins des revendications 4 à 6, dans laquelle le second canal (140) est un second alésage oblique ou perpendiculaire à la direction de déplacement du piston de commande (120), lequel alésage perce une paroi latérale du piston de commande (120).
  8. Soupape à siège (100) suivant l'une au moins des revendications 4 à 7, dans laquelle le piston de commande (120) est précontraint par un ressort (110) dans la direction de fermeture.
  9. Soupape à siège (100) suivant la revendication 8, dans laquelle la force du ressort (110) est conçue de sorte qu'elle compense des forces de frottement.
  10. Soupape à siège (100) suivant l'une au moins des revendications 1 à 9, dans laquelle la première, la deuxième et la troisième surfaces de sollicitation (F1, F2, F3) sont associées à un premier, deuxième et troisième compartiments de pression (F1A, F2A, F3A) correspondants à l'intérieur de la soupape à siège (100), le premier et le troisième compartiments de pression (F1A, F3A) étant en liaison fluidique mutuelle par le premier canal (130).
  11. Soupape à siège (100) suivant la revendication 10, dans laquelle le deuxième compartiment de pression (F2A) est disposé dans la soupape à siège (100) de sorte que le second canal (140) se déplace sur le deuxième compartiment de pression (F2A), lorsque le passage entre l'entrée (E) et la sortie (A) s'ouvre.
  12. Soupape à siège (100) suivant la revendication 11, dans laquelle le second canal (140) est disposé dans le piston de commande (120) de sorte qu'il crée une liaison entre le deuxième compartiment de pression (F2A) et le premier ou le troisième compartiment de pression (F1A, F3A), lorsque le piston de commande (120) atteint au moins 70% de la course maximale dans la direction d'ouverture, le passage entre l'entrée (E) et la sortie (A) étant totalement ouvert en position de course maximale.
EP09005478A 2009-04-17 2009-04-17 Soupape à siège dotée d'une fonction de soupape de dérivation et de compensateur de pression Not-in-force EP2241764B1 (fr)

Priority Applications (2)

Application Number Priority Date Filing Date Title
EP09005478A EP2241764B1 (fr) 2009-04-17 2009-04-17 Soupape à siège dotée d'une fonction de soupape de dérivation et de compensateur de pression
AT09005478T ATE522728T1 (de) 2009-04-17 2009-04-17 Sitzventil mit umlaufventil- und druckwaagefunktion

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP09005478A EP2241764B1 (fr) 2009-04-17 2009-04-17 Soupape à siège dotée d'une fonction de soupape de dérivation et de compensateur de pression

Publications (2)

Publication Number Publication Date
EP2241764A1 EP2241764A1 (fr) 2010-10-20
EP2241764B1 true EP2241764B1 (fr) 2011-08-31

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DE102011121052A1 (de) * 2011-12-14 2013-06-20 Robert Bosch Gmbh Druckbegrenzungsventil und geschlossenerhydraulischer Kreis mit einem Druckbegrenzungsventil
DE102021001960A1 (de) 2021-04-14 2022-10-20 Hydac Fluidtechnik Gmbh Ventil

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US5137254A (en) * 1991-09-03 1992-08-11 Caterpillar Inc. Pressure compensated flow amplifying poppet valve
DE29810860U1 (de) 1998-06-17 1998-08-13 Heilmeier & Weinlein Fabrik für Oel-Hydraulik GmbH & Co KG, 81673 München Hydraulische Steuervorrichtung
US7621211B2 (en) * 2007-05-31 2009-11-24 Caterpillar Inc. Force feedback poppet valve having an integrated pressure compensator

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EP2241764A1 (fr) 2010-10-20

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