EP1694965B1 - Dispositif de regulation de puissance totale - Google Patents

Dispositif de regulation de puissance totale Download PDF

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Publication number
EP1694965B1
EP1694965B1 EP04797833A EP04797833A EP1694965B1 EP 1694965 B1 EP1694965 B1 EP 1694965B1 EP 04797833 A EP04797833 A EP 04797833A EP 04797833 A EP04797833 A EP 04797833A EP 1694965 B1 EP1694965 B1 EP 1694965B1
Authority
EP
European Patent Office
Prior art keywords
valve
pressure
total power
piston
power control
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Fee Related
Application number
EP04797833A
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German (de)
English (en)
Other versions
EP1694965A1 (fr
Inventor
Juan Moya
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.)
Brueninghaus Hydromatik GmbH
Original Assignee
Brueninghaus Hydromatik GmbH
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Brueninghaus Hydromatik GmbH filed Critical Brueninghaus Hydromatik GmbH
Publication of EP1694965A1 publication Critical patent/EP1694965A1/fr
Application granted granted Critical
Publication of EP1694965B1 publication Critical patent/EP1694965B1/fr
Expired - Fee Related legal-status Critical Current
Anticipated expiration legal-status Critical

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/002—Hydraulic systems to change the pump delivery
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2201/00—Pump parameters
    • F04B2201/12—Parameters of driving or driven means
    • F04B2201/1203—Power on the axis
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2203/00—Motor parameters
    • F04B2203/02—Motor parameters of rotating electric motors
    • F04B2203/0208—Power
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B2203/00—Motor parameters
    • F04B2203/06—Motor parameters of internal combustion engines
    • F04B2203/0604—Power

Definitions

  • the invention relates to a total power control device for two pumps, according to the preamble of claim 1.
  • Such a buzzer power control device is known from EP 0 561 153 B1.
  • the second leg of the reversing lever acts on a power control valve, by which a setting pressure is adjusted, which acts on a connected to an adjusting mechanism of the pump actuating piston.
  • a counterforce is generated on the first leg of the reversing lever, which depends on the working pressure of the second hydraulic pump.
  • the counterforce decreases with increasing pressure, which is generated by the second hydraulic pump from.
  • a cylinder is used, in which a piston is arranged, which is acted upon by an adjustable spring in the direction of the first leg of the reversing lever.
  • the guided in the cylinder piston has a piston surface which is acted upon by the working pressure of the second hydraulic pump. The increasing with the working pressure hydraulic force acts against the force of the adjustable spring and acting on the leg of the bellcrank counterforce is reduced.
  • the second hydraulic pump is constructed completely identical to the first hydraulic pump described above, wherein a cylinder is likewise provided, in which a piston is arranged for adjusting the counterforce acting on the deflection lever of the second hydraulic pump. According to the first hydraulic pump, the piston surface of the piston is now acted upon by the working pressure of the first hydraulic pump to generate the counterforce for the second hydraulic pump.
  • the described control device for controlling the total power has the disadvantage that in each case a cylinder with a piston arranged therein must be provided for both hydraulic pumps, which still acts on the reversing lever. In order to achieve a symmetrical setting, the pressure spring acting on the piston must be precisely adjusted.
  • Another disadvantage is the space required by the cylinder to generate the counterforce. This contradicts in particular the endeavor to create a system that is as simple and compact as possible primary drive source is provided for driving two hydraulic pumps.
  • the invention is therefore based on the object to provide a total power control device, which takes into account the power absorbed by the other pump in the power control, without requiring additional components compared to a separate power control.
  • the delivery volume of two pumps is separately adjustable.
  • the delivery volume is changed by a respective adjusting device connected to one of the pumps and thus sets the volume flow delivered in each case to one working line.
  • To actuate the adjusting acts in the adjusting a control pressure that is adjustable by a sum power control valve.
  • each adjusting device is associated with a total power control valve, in each of which a measuring surface is arranged, which is acted upon by the working pressure supplied by the respective other pump in the working line connected to it.
  • the working pressure generated by the other pump is thus taken into account when setting the setting pressure.
  • the working pressure of the other pump is used as a measure of the power consumed by the second pump.
  • the summation power control valves of the hydraulic pumps are designed as valve cartridges.
  • a simple conversion from a pump with a conventional power control valve to a total power control is possible by the different valve cartridges are replaced with each other.
  • the measuring surface on the valve piston of the valve cartridge is designed as an annular surface.
  • the design of the measuring surface as an annular surface makes it possible to use the free end face of the valve piston for further introduction of force, for example by a lever in a hyperbola regulator.
  • Another advantage is that the measuring surface is arranged in the axial direction between two non-pressurized areas.
  • the small, but unavoidable leakage flow which arises when the annular surface is acted upon by the working pressure of the other hydraulic pump, can be dissipated in a simple manner. This is preferably achieved by providing a volume connected to a tank connection adjacent to the space in which the annular surface is arranged.
  • a spring space is provided on the other side adjacent to the space in which the annular surface is arranged, the volume of which is also connected to the tank volume.
  • this spring chamber at least one compression spring is arranged, which acts on the valve piston with a force axially, against which the valve piston is acted upon both by the working pressure of the other hydraulic pump and by the power of the hydraulic pump to be adjusted proportional force.
  • the power proportional to the power acts for this purpose on the side facing away from the spring chamber end of the valve piston.
  • Fig. 1 the inventive sum power control device for a first hydraulic pump unit 1 and a second hydraulic pump unit 41 is shown, which is shown in the lower part of Fig. 1.
  • the first hydraulic pump unit 1 and the second hydraulic pump unit 41 are comparable in terms of their structure, which is why the detailed description of the individual elements and their function is based only on the first hydraulic pump unit 1.
  • the first hydraulic pump unit 1 has a pump 2, which is driven via a drive shaft 3 by a primary drive machine, not shown.
  • a primary drive machine such an engine may be, for example, a diesel engine or an electric motor.
  • the pump 2 is provided for conveying in a first hydraulic circuit and sucks this via a suction line 4 pressure medium and promotes it in a working line 5.
  • the provided in the illustrated embodiment pump 2 is designed only for conveying in only one direction, since it is in the exemplary embodiment is an open circuit. However, the invention can also be used in closed circuits.
  • the pump 2 can be adjusted in its delivery volume.
  • the adjustment is made by an adjusting device 6.
  • the adjusting device 6 comprises a cylinder 7, in which a longitudinally displaceable Control piston 8 is arranged. This adjusting piston 8 is connected to the pump 2 for adjusting its delivery volume via a linkage 9.
  • the actuating piston 8 has a first piston surface 8 'and a second piston surface 8' ', which are oriented opposite to each other and which can be acted upon in a working pressure chamber 10 and a control pressure chamber 12 with a force.
  • the first piston surface 8 ' is smaller than the second piston surface 8' ', wherein the force acting on the first piston surface 8' hydraulic force is supported by a spring 11 which acts on the actuating piston 8 in the direction of the control pressure chamber 12 with a force.
  • a displacement of the control piston 8 in the direction of the control pressure chamber 12 causes an adjustment of the pump 2 in the direction of its maximum delivery volume.
  • the pumping means conveys the pressure medium into the working line 5
  • the pressure generated in the working pressure chamber 10 is exactly this pressure generated in the working line 5.
  • the working line 5 is connected to the working pressure chamber 10 via a working pressure supply line 13 and a first branch 14 branching off from it.
  • a pressure which is proportional to the pressure prevailing in the working line 5 working pressure.
  • This working pressure acts on the first piston surface 8 'of the adjusting piston 8 and acts on it together with the force of the spring 11 so that the pump 2 is adjusted in the direction of its maximum delivery volume.
  • a defined actuating pressure is set in the actuating pressure chamber 12. If an equilibrium is thus established between the forces acting in the working pressure chamber 10 and in the control pressure chamber 12 on the control piston 8, there is no further adjustment of the delivery volume of the pump 2.
  • the control pressure chamber 12 is connected via a control pressure channel 15 , a pressure regulating valve 16 and a connecting passage 17 are connected to a cumulative power control valve 18.
  • the connecting channel 17 connects the total power control valve 18 to the pressure regulating valve 16, which in its rest position represents an unthrottled connection between the connecting duct 17 and the actuating pressure line 15.
  • the total power control valve 18 is a 3/2-way valve, which is connected on the input side to a connection channel 20 and a tank channel 21.
  • the connection channel 20 leads to the total power control valve 18 to the prevailing in the working line 5 working pressure.
  • the connection channel 20 is connected to a second branch 19, which in turn is connected to the first working pressure supply line 13.
  • the position of the total power control valve 18 is determined by an adjustable compression spring 22 and the forces acting counter to the adjustable compression spring 22 on a plunger 23 and a measuring surface 24.
  • a plunger 23 acts on the measuring surface 24 of the pressure generated in the working line of the second hydraulic pump unit 41 and generates a hydraulic force.
  • the plunger 23 acts via a lever 25 which is rotatably mounted about a pivot point 26, a force which is proportional to the absorbed power of the pump 2.
  • the total power control valve 18 In its rest position, ie when the second hydraulic pump unit 41 generates no working pressure and also the first hydraulic pump unit 1 does not absorb any power, the total power control valve 18 is held by the adjustable compression spring 22 in its first end position shown in FIG. In the first end position of the cumulative power control valve 18, the connecting channel 17 is connected via the tank channel 21 with a tank volume 27. Thus, the control pressure chamber 12 is relaxed by the total power control valve 18 via the control pressure channel 15, the through-connected pressure control valve 16 and the connecting channel 17 and finally via the tank channel 21 into the tank volume 27.
  • the falling control pressure in the control pressure chamber 12 has a movement of the control piston 8 in Fig. 1 to the right due to the first prevailing in the working pressure chamber 10 unchanged pressure. Thus, the pump 2 is moved in the direction of larger delivery volume via the linkage 9.
  • the adjustment of the delivery volume of the pump 2 follows the hyperbolic curve of the performance curve. In the direction of greater working pressures, this characteristic asymptotically approaches a corresponding minimum delivery volume. However, this is associated with a strong increase in pressure. In order to prevent this, and thus to ensure that a permissible maximum pressure is not exceeded in the line system, the adjusting pressure chamber 12 is depressed by the pressure regulating valve 16 above this maximum maximum pressure and thus the pump 2 is adjusted in the direction of smaller delivery volume. In this case, the power control is overridden by the pressure control valve 16.
  • the connecting channel 17 is connected unthrottled to the actuating pressure channel 15.
  • the pressure relief valve 16 is held by a further adjustable compression spring 32 in this position.
  • a delivery pressure measuring surface 33 is acted upon by the pressure prevailing in the working line 5.
  • the delivery pressure measuring surface 33 is oriented in such a way that the hydraulic force acting on the pressure regulating valve 16 or its valve piston is directed against the force of the further adjustable compression spring 32.
  • a certain pressure limit is exceeded by the pressure in the working line 5 is thus against the force of the adjustable compression spring 32, which consequently determines this pressure limit, the pressure control valve 16 is adjusted in the direction of its second end position and the control pressure chamber 12 is depressed with the prevailing in the working line 5 working pressure , As a result, the actuating piston 8 is moved in Fig. 1 to the left and the pump 2 is adjusted in the direction of smaller delivery volume.
  • a measuring channel 34 is provided, in which a throttle 35 is arranged.
  • a throttled connection 38 is provided, which connects the connecting channel 17, bypassing the pressure regulating valve 16 with the actuating pressure channel 15.
  • the throttled connection 38 is preferably led out of the housing of the first hydraulic pump unit 1 and can be tapped off at a measuring connection 39.
  • a further hydraulic pump unit 41 is driven by the same primary drive machine, then during adjustment the power of the first hydraulic pump unit 1, the power absorbed by the second hydraulic pump unit 41 are taken into account. This is done via the measuring surface 24, which is formed on the summation power control valve 18. The measuring surface 24 is acted upon by a pressure and thus generates a hydraulic force which acts in the same direction with the force acting on the plunger 23 against the force of the adjustable compression spring 22.
  • the measuring surface 24 is connected to the second hydraulic pump unit 41 via a first connecting line 36.
  • the corresponding elements of the second hydraulic pump unit 41 are provided with reference numerals, which are increased by 40 relative to the reference number of the corresponding element of the first hydraulic pump unit 1.
  • the first connecting line 36 is connected to the working pressure line 53 of the second hydraulic pump unit 41.
  • the working pressure generated in the working line 45 of the second hydraulic pump unit 41 by the pump 42 of the second hydraulic motor unit 41 is supplied via the working pressure supply line 53 of the second hydraulic pump unit 41 and the first connecting line 36 to the measuring surface 24 of the total power control valve 18 of the first hydraulic pump unit 1.
  • a second connecting line 37 is provided, through which the working pressure prevailing in the working line 5 and thus the working pressure supply line 13 of the first hydraulic pump unit 1 working pressure of the measuring surface 64 of the cumulative power control valve 58 of the second hydraulic pump unit 41 becomes.
  • the power consumed by the first hydraulic pump unit 1 is also taken into account when setting the control pressure for the adjusting device 46 of the second hydraulic pump unit 41.
  • the total power control valves 18 and 58 are used as so-called valve cartridges in the housings of the hydraulic pump units.
  • the respective housing by the dash-dotted line are schematically shown, which surrounds all located within the housing elements and which is designated by the reference numeral 1.
  • the pressure control valves 16 and 56 are preferably designed as valve cartridges and are inserted into a corresponding bore in the housing of the respective hydraulic pump unit 1 or 41.
  • FIG. 1 A preferred embodiment of such a valve cartridge 81 of a cumulative power control valve 18 and 58 of the sum power control device according to the invention is shown in FIG.
  • the valve cartridge 81 is inserted into a designated opening of the housing of the first hydraulic pump unit 1 and the second hydraulic pump unit 41.
  • a Thread provided, which is screwed into a corresponding thread of the housing of the hydraulic pump unit and is sealed by means of a sealing ring 83.
  • a valve sleeve 84 connects on the projecting into the housing of the hydraulic pump unit side of the valve housing 82 in the axial direction.
  • valve sleeve 84 is penetrated axially by a stepped recess into which a valve piston 85 is inserted.
  • This valve piston 85 has at one end an extension 86 which projects slightly out of the valve sleeve 84 in the direction of the valve housing 82.
  • the valve housing 82 also has a blind hole designed as a central recess into which a first spring 87 and a second spring 88 are used.
  • the first spring 87 and the second spring 88 are designed as compression springs and are received by the spring chamber 89 forming a central recess of the valve housing 82.
  • the first spring and the second spring 87 and 88 are each supported on a first spring seat 90 and a second spring seat 91.
  • the first spring seat 90 has in the middle an axially extending centering for the first spring 87 and the second spring 88, which is penetrated by a longitudinal bore 92.
  • the first spring seat 90 has a substantially disk-shaped geometry, which has a recess on the side facing away from the centering, in which the extension 86 of the valve piston 85 engages, so that pressure forces are transmitted between the valve piston 85 and the first spring seat 90 in the axial direction can.
  • the second spring seat 91 is arranged, which in turn has a recess on one side and a centering for centering the first spring 87 and the second spring 88 on the opposite side of the depression.
  • an adjusting screw 93 a In the recess of the second spring seat 91 engages one end an adjusting screw 93 a.
  • the adjusting screw 93 is screwed into a thread arranged in the valve housing 82, so that by further screwing the adjusting screw 93, the distance between the first spring seat 90 and the second spring seat 91 can be reduced.
  • the voltage of the first spring 87 and the second spring 88 can be changed, thereby adjusting the characteristic of the cumulative power control valve 18 or 58.
  • the adjusting screw 93 is countered by means of a lock nut 94 against the valve housing 82.
  • Another protective measure is the screwing on a threaded cap 95, which prevents contamination or corrosion of the adjusting screw 93.
  • the extension 86 is arranged at one end of the valve piston 85 frontally and executed approximately dome-shaped. On the opposite end face 96 of the valve piston 85, however, a recess 97 is formed. This recess 97 serves to receive the known from Fig. 1 plunger 23 and may be provided to fix it with an internal thread.
  • the valve piston 85 has a first guide section 98, axially thereof a second guide section 99 and a third guide section 100 which is again arranged at an axial, greater distance therefrom.
  • This third guide section 100 is arranged axially in the region of the recess 97 and has a preferably identical diameter as the second guide section 99.
  • the first guide section 98 is enlarged in its diameter.
  • This radial extension of the valve piston 85 generates at the oriented in the direction of the end face 96 end of the first guide portion 98 an annular surface 101, the measuring surface 24 and 64 of the first hydraulic pump unit 1 and the second hydraulic pump unit 41 of FIG. 1 corresponds.
  • the continuous recess of the valve sleeve 84 has a radial step 102 corresponding to the different diameters of the first guide section 98 and the second and third guide sections 99 and 100.
  • This radial step 102 is arranged corresponding to the distance between the first guide section 98 and the second guide section 99 axially offset from the annular surface 101, so that between the annular surface 101 and the radial step 102, an annular space 103 is formed.
  • This annular space 103 is connected via radially arranged first holes 104 with a arranged on the outside of the valve sleeve 84 circumferential first groove 105.
  • this circumferential first groove 105 opens, on the part of the first hydraulic pump unit 1, for example, the first connecting line 36, as is merely indicated in FIG.
  • the first guide portion 98 and the second guide portion 99 sealingly cooperate with the corresponding portions of the valve sleeve 84.
  • the annular space z. B. be acted upon via the first connecting line 36 with a pressure which generates on the annular surface 101, a hydraulic force in the axial direction against the force of the first spring 87 and the second spring 88.
  • a radially tapered section 106 adjoins the second guide section 99, whereby in this region of the valve piston 85 again an annular space is created, into which second bores 107 arranged radially in the valve sleeve 84 open.
  • These second holes 107 connect the annular portion formed around the radially tapered portion 106 with a circumferential second groove 108 disposed on the circumference of the valve sleeve 84.
  • the first radially tapered portion 106 extends to a first control edge 111, which is formed by a renewed radial expansion of the valve piston 85.
  • first control edge 111 which is formed by a renewed radial expansion of the valve piston 85.
  • third holes 109 which are arranged radially in the valve sleeve 84 and open out into a circumferential third groove 110, are just covered by the first control edge 111, so that between the third holes 109 and the second holes 107 no pressure medium flow is possible.
  • a second control edge 115 is further formed on the valve piston 85 by a radial step, to which a second radially tapered portion 112 connects, which extends to the third guide portion 100.
  • the second control edge 115 is again arranged so that in a middle position of the valve piston 85, a connection of the third holes 109 and to the arranged in the region of the second radially tapered portion 112 fourth holes 113 is not made. In this position of the valve piston 85 is thus from the third holes 109, neither to the second holes 107 still to the fourth holes 113 a flow-through cross-section available. In this state of equilibrium so that the control pressure in the control pressure chamber 12 is not changed and the set displacement remains constant.
  • valve piston 85 If, on the other hand, the valve piston 85 is displaced against the force of the first spring 87 and the second spring 88 by a hydraulic force on the annular surface 101 or a larger force acting on the end face 96 of the valve piston 85, the second control edge 115 releases a cross-section through which over which the third bores 109 and the fourth bores 113 are connected to one another.
  • the arrangement of the bores in the axial direction is preferably carried out as shown in the embodiment of FIG. 2, that is, so that the first holes 104 through which the annular space 103 and thus the annular surface 101 is acted upon by the working pressure of the respective other hydraulic pump unit is, between the spring chamber 89 and the second holes 107 are arranged. Since the second holes 107 are connected via the tank channel 21 with the tank volume 27 of the hydraulic pump unit and the spring chamber 89 is depressurized, both a leakage path of the pressure medium from the annular space 103 past the first guide portion 98 and the second guide portion 99 over given, wherein the escaping leakage fluid is discharged in each case via an adjacent pressureless volume in the tank volume 27.
  • the spring chamber 89 is also coupled via a drain hole 116 to the tank volume 27.
  • the invention is not limited to the embodiment described and applicable, for example, in a closed circuit. Furthermore, all described or drawn features can be combined with each other.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Control Of Positive-Displacement Pumps (AREA)
  • Details Of Reciprocating Pumps (AREA)
  • Fluid-Pressure Circuits (AREA)

Claims (6)

  1. Dispositif de régulation de puissance totale pour au moins deux pompes (2, 42) qui sont chacune raccordées à une conduite de travail (5, 45) et dont le débit est réglable séparément par un dispositif de réglage (6, 46) respectif, une pression de réglage agissant sur chaque dispositif de réglage étant réglable par une soupape de régulation de puissance totale (18, 58), caractérisé en ce que chaque soupape de régulation de puissance totale (18, 58) présente une surface de mesure (24, 64) sur un piston de soupape, la surface de mesure (24, 64) de la soupape de régulation de puissance totale (18, 58) d' pompe (2, 42) étant soumise directement à une pression de travail de l'autre pompe (42, 2) et en ce que le piston de soupape (85) de la soupape de régulation de puissance totale (18, 58) d'une pompe (2, 42) peut être soumis à une force proportionnelle à la puissance de cette pompe (2, 42) de même sens que la force hydraulique s'exerçant sur la surface de mesure.
  2. Dispositif de régulation de puissance totale selon la revendication 1, caractérisé en ce que les soupapes de régulation de puissance totale (18, 58) sont réalisées sous la forme de cartouches de soupapes (81).
  3. Dispositif de régulation de puissance totale selon la revendication 1 ou 2, caractérisé en ce que sur le piston de soupape (85) est formée une surface annulaire (101) qui forme la surface de mesure (24, 64).
  4. Dispositif de régulation de puissance totale selon la revendication 3, caractérisé en ce que la surface annulaire (101) est réalisée de telle manière qu'elle est disposée dans la cartouche de soupape (81) dans la direction axiale entre deux espaces (89) reliés à un volume de réservoir (27).
  5. Dispositif de régulation de puissance totale selon l'une des revendications 1 à 4, caractérisé en ce que la force hydraulique s'exerçant sur la surface de mesure (24, 64) et la force proportionnelle à la puissance s'exercent sur le piston de soupape (85) en s'opposant à un ressort (87, 88) s'appuyant contre une extrémité côté frontal.
  6. Dispositif de régulation de puissance totale selon l'une des revendication 1 à 5, caractérisé en ce que la surface de mesure (24, 64) de la soupape de régulation de puissance totale (18, 58) d'une pompe (2, 42), pour amener la pression de travail à l'autre pompe (42, 2), est reliée à une conduite de travail (45, 5) de l'autre pompe (42, 2) par l'intermédiaire d'une conduite de liaison (36, 37).
EP04797833A 2003-12-15 2004-11-11 Dispositif de regulation de puissance totale Expired - Fee Related EP1694965B1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE10358727A DE10358727B3 (de) 2003-12-15 2003-12-15 Summenleistungsregelvorrichtung
PCT/EP2004/012808 WO2005057011A1 (fr) 2003-12-15 2004-11-11 Dispositif de regulation de puissance totale

Publications (2)

Publication Number Publication Date
EP1694965A1 EP1694965A1 (fr) 2006-08-30
EP1694965B1 true EP1694965B1 (fr) 2007-03-28

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ID=34672759

Family Applications (1)

Application Number Title Priority Date Filing Date
EP04797833A Expired - Fee Related EP1694965B1 (fr) 2003-12-15 2004-11-11 Dispositif de regulation de puissance totale

Country Status (4)

Country Link
US (1) US7607297B2 (fr)
EP (1) EP1694965B1 (fr)
DE (2) DE10358727B3 (fr)
WO (1) WO2005057011A1 (fr)

Families Citing this family (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102007044451A1 (de) * 2007-09-18 2009-03-19 Robert Bosch Gmbh Anschlussplatte für eine hydrostatische Kolbenmaschine
DE102010048068B4 (de) * 2010-04-16 2022-11-10 Robert Bosch Gmbh Ventilanordnung
DE102017112700A1 (de) * 2017-06-08 2018-12-13 Schwäbische Hüttenwerke Automotive GmbH Steuerventil

Family Cites Families (8)

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Publication number Priority date Publication date Assignee Title
DE3638889A1 (de) * 1986-11-14 1988-05-26 Hydromatik Gmbh Summen-leistungsregelvorrichtung fuer wenigstens zwei hydrostatische getriebe
DE3919175A1 (de) 1989-06-12 1990-12-13 Hydromatik Gmbh Regeleinrichtung fuer eine verstellbare pumpe
DE4208925C1 (fr) * 1992-03-19 1993-06-03 Hydromatik Gmbh, 7915 Elchingen, De
US5567123A (en) * 1995-09-12 1996-10-22 Caterpillar Inc. Pump displacement control for a variable displacement pump
JP3891616B2 (ja) * 1996-11-14 2007-03-14 株式会社加藤製作所 可変容量ポンプの入力トルク制御回路
JP3549989B2 (ja) * 1996-12-10 2004-08-04 日立建機株式会社 油圧作業機の油圧回路装置
JP3383754B2 (ja) * 1997-09-29 2003-03-04 日立建機株式会社 油圧建設機械の油圧ポンプのトルク制御装置
JP2001254681A (ja) * 2000-03-13 2001-09-21 Kato Works Co Ltd 可変容量ポンプの入力トルク制御回路

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
None *

Also Published As

Publication number Publication date
US7607297B2 (en) 2009-10-27
EP1694965A1 (fr) 2006-08-30
DE502004003366D1 (de) 2007-05-10
WO2005057011A1 (fr) 2005-06-23
DE10358727B3 (de) 2005-08-25
US20070151236A1 (en) 2007-07-05

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