EP0211096B1 - Réglage de puissance ou limitation de couple - Google Patents

Réglage de puissance ou limitation de couple Download PDF

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Publication number
EP0211096B1
EP0211096B1 EP19850109781 EP85109781A EP0211096B1 EP 0211096 B1 EP0211096 B1 EP 0211096B1 EP 19850109781 EP19850109781 EP 19850109781 EP 85109781 A EP85109781 A EP 85109781A EP 0211096 B1 EP0211096 B1 EP 0211096B1
Authority
EP
European Patent Office
Prior art keywords
control
pressure
piston
power
spring
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
EP19850109781
Other languages
German (de)
English (en)
Other versions
EP0211096A1 (fr
Inventor
David A. Jacombs
Gerald Warren
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.)
Vickers Systems GmbH
Original Assignee
Vickers Systems 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 Vickers Systems GmbH filed Critical Vickers Systems GmbH
Priority to EP19850109781 priority Critical patent/EP0211096B1/fr
Priority to DE8585109781T priority patent/DE3578196D1/de
Publication of EP0211096A1 publication Critical patent/EP0211096A1/fr
Application granted granted Critical
Publication of EP0211096B1 publication Critical patent/EP0211096B1/fr
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

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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
    • F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
    • F04B1/12—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinder axes coaxial with, or parallel or inclined to, main shaft axis
    • F04B1/26—Control
    • F04B1/30—Control of machines or pumps with rotary cylinder blocks
    • F04B1/32—Control of machines or pumps with rotary cylinder blocks by varying the relative positions of a swash plate and a cylinder block
    • F04B1/324—Control of machines or pumps with rotary cylinder blocks by varying the relative positions of a swash plate and a cylinder block by changing the inclination of the swash plate

Definitions

  • the invention relates to a power-controlled hydraulic pump with a variable displacement (control pump).
  • Power-controlled control pumps are known and are used, among other things, when the drive motor (electric motor or internal combustion engine) should have a smaller output than the so-called basic performance value, which is calculated from the largest displacement at the highest permissible pressure. If one assumes a drive motor with an approximately constant output speed, then power-controlled actuating pumps also have the torque limitation feature.
  • the output of a control pump is calculated from the pump current multiplied by the working pressure.
  • the pump current and the working pressure have to be measured or corresponding values have to be formed and processed. While the working pressure can be used directly on the piston surfaces of a hydraulic control device, measuring the pump current requires a greater effort, which essentially determines the cost of the power control.
  • a hydraulic pump with variable displacement volume is already known, in which a control valve is provided with opposing first and second control piston surfaces, of which the first, larger control piston surface is acted upon by the working pressure and acts against the force of the valve spring, while the second , smaller control piston area is acted on by the control pressure.
  • the control valve is designed to switch and has a middle closed position, in which the control channel is neither connected to the working pressure nor to the outlet to the tank.
  • the control valve is used to regulate the displacement volume, but it does not appear from the document that this should be done with a view to a certain maximum power limitation.
  • the same control valve can also be used to regulate a double pump (US Pat. No. 3,093,081), the two pumps being different in size and successively adjusted in terms of their displacement volume. It does not appear from the writing that this should be done to limit performance.
  • control in particular for overload protection of the drive motor, can be achieved with significantly reduced technical outlay. It is particularly advantageous that the control valve does not have to be attached to the control pump itself.
  • the piston return device is designed so that each position of the actuating piston corresponds to a certain level of the control pressure.
  • This requirement is met for a spring or a spring assembly as a piston return device, regardless of whether there is direct action on the actuating piston or via an intermediate element, for example the swash plate of an axial piston pump.
  • the piston adjustment device should have a non-linear characteristic, with decreasing pump current e.g. increasing spring stiffness. The control pressure should therefore increase more than it corresponds to the decrease in the pump current.
  • a simple control valve with two piston surfaces is sufficient to carry out the invention, the one, smaller, piston surface being subjected to the working pressure and the other, larger, piston surface being subjected to the control pressure.
  • the control valve forms, so to speak, a product of the working pressure and the pump current, the size of the pump current being entered into the power control as a pressure signal in the form of the control pressure.
  • the control valve can be designed as a spring-loaded slide piston with two control edges. This not only represents an extremely economical solution to the performance problem, it is also extremely small in size (hardly larger than a conventional compensator with two slide pistons) and can be accommodated in the valve housing by other valves.
  • the axial piston pump 1 has a pump housing 2 with a housing cover 3, which is designed as a control plate, a swash plate 4, a cylinder drum 5 with a pump piston 6 mounted therein and an actuating piston 7 are also provided.
  • the cylinder drum 5 has splines for driving by a drive shaft 8, which is mounted in bearings in the housing 2 or the housing cover 3.
  • the pistons 6 are supported on the swash plate 4 via sliding shoes 9.
  • the swash plate 4 is pivotally mounted and is pushed via a return device 10 in the direction of the maximum inclined position against the possibility of the actuating piston 7 acting.
  • inlet and outlet channels 11, 12 are provided, each ending in kidney-shaped slots on the inside of the cover 3.
  • Hydraulic fluid is sucked in from a tank 13 via the inlet channel 11, moved to the pressure side by rotation of the cylinder drum 5 and displaced into the outlet channel 12.
  • a corresponding working pressure or system pressure builds up in the outlet channel.
  • Fig. 2 shows the return device 10 and the actuating piston 7 in an enlarged view.
  • the return device contains a spring assembly with two compression springs 15 and 16 which are fitted into one another and whose spring travel is of different sizes.
  • the spring 16 represents an additional spring which only comes into effect after the swash plate 4 has been pivoted to a certain extent and then exerts a force in addition to the spring 15.
  • a kinked spring characteristic of the overall arrangement is achieved in this way. Desirable is a spring characteristic that becomes progressively steeper in a force-displacement diagram.
  • the actuating piston 7 is guided in a cylinder 17 which is connected to a control valve 20 via a control line 18. This is connected to the pump outlet channel 12 via a pressure line 19. A drain line 14 connects to the tank 13.
  • the control valve 20 is accommodated in a housing 21 which can also accommodate further valves, for example a pressure relief valve 40.
  • the control valve 20 has a slide piston 22 which is guided in a bore 23 with the cross-sectional area A1 and has a control piston collar 24 with two control edges, which cooperate with corresponding edges between the valve bore 23 and a control bore 25.
  • the valve bore 23 is connected to the working pressure via the pressure line 19, while there is a connection to the drain line 14 to the right of the control piston collar 24.
  • this connection runs via a control channel 35 and a valve bore 33 of the pressure relief valve 40, which has a slide piston 32 and a control piston collar 44, which normally shuts off the system pressure supplied via 19 from the control channel 35.
  • a throttled leakage current flows from the pressure line 19 to the drain line 14, so that an average pressure is established in the control bore 25, which pressure is supplied to the actuating piston 7 as a control pressure via the control pressure line 18.
  • the control device operates as follows: When starting, the swash plate 4 is in its maximum swivel position, i.e. the axial piston pump is set to the maximum displacement. It is assumed that the drive motor, not shown, which engages the shaft 8 is brought to the desired speed. The maximum pump current Q1 (FIGS. 3 and 4) is then supplied to a hydraulic consumer, not shown, the properties of which determine the level of the working pressure which is established in the outlet line 12. It is assumed that a higher pressure than P1 (FIGS. 3 and 4) arises. An imbalance is reached on the spool 22, i.e.
  • the spool moves to the right in the drawing and releases a larger throttle cross section between its pressure supply side and the control bore 25, so that the control pressure in the control pressure line 18 increases and the actuating piston 7 is displaced.
  • the swash plate 4 is pivoted into a position for reduced pump current, which is represented by the falling characteristic curve in the QP diagram of FIGS. 3 and 4.
  • the increased control pressure is fed back to the piston 27, so that the valve 20 moves back into its equilibrium position, the filling of the cylinder 17 essentially not changing any further.
  • the pump then delivers a certain pump current at the requested working pressure. If a further increased working pressure is requested by the consumer, this leads to the control valve 20 responding again and the control pressure being increased further, with the result that the swash plate 4 is adjusted to a further reduced pump delivery flow.
  • the inclination of the QP characteristic in Fig. 3 or 4 is determined by the ratio of the cross-sectional areas of the control collar 24 to the control piston 27, i.e. determined according to the value A1 / A2.
  • the behavior of the control device can accordingly be determined by appropriate selection of the ratio A1: A2.
  • Another means of influencing the control characteristic is to influence the characteristic of the return device 10.
  • this is done by the spring 16, which may have a spring characteristic K2. It is assumed that the spring 16 comes into effect at the operating point Q2 P2 (FIG. 4), that is to say is acted upon as a result of the retracting swash plate 4.
  • the further swiveling of the swash plate 4 on a reduced pump flow is thus opposed by the two springs 15 and 16 with the characteristic curves K1 and K2. It therefore becomes a relative requires greater actuating piston force, which can only be generated by a relatively higher control pressure P c . According to the equation given, this leads to a higher working or system pressure P s compared to the case without the spring 16.
  • the "broken" characteristic according to FIG. 4 is therefore achieved.
  • the pressure relief valve 40 responds and increases the control pressure Pc to such an extent that the pump swivels into its zero stroke position. This is shown at P3 in FIG. 4.
  • the slope of the characteristic between P1 and P2 in FIG. 3 represents a rough approximation to a hyperbola.
  • the approximation to a hyperbola is better in the characteristic of the FIG.
  • the approach to a hyperbola can be driven further by appropriate design of the return device 10. In general, however, it is sufficient to achieve the approximation according to FIG. 3 or 4 to the ideal hyperbola when it comes to creating only overload protection for the drive motor.
  • the overload protection control valve 20 can be coupled to further valves in order to regulate the servomotor, because it only responds in the limit case for overload protection purposes.
  • Fig. 5 shows an axial piston pump 1a, in which two actuating pistons 7a, 7b are provided. While the - larger - control piston 7a is connected to the overload protection control valve 20 via the control line 18, the - smaller - control piston 7b is connected to a branch line 12b of the outlet line 12 and receives the system pressure.
  • the smaller actuating piston 7b forms, together with a return spring (not shown), the return device 10 of the swash plate 4.
  • an increasing control pressure in the cylinder 17 can be achieved as a function of the swivel position of the swash plate 4.

Landscapes

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

Claims (4)

1. Pompe hydraulique à régulation de puissance et à volume de refoulement variable, possédant les caractéristiques suivantes :
pour la variation du volume de refoulement (Q), il est prévu un organe de réglage (4, 7, 10) comportant un piston de commande (7) ; le piston de commande (7) peut être déplacé par une pression de commande (Pc) à l'encontre de l'action
d'un dispositif de rappel (10) ; le dispositif de rappel (10) est constitué de manière qu'à chaque position du piston de commande (7), corresponde une valeur déterminée de la pression de commande (Pc) ;
un distributeur (20) règle la valeur de la pression de commande (Pc) et comprend à cet effet un organe de commande (22) qui présente une première surface de piston de commande (A1) qui est chargée par la pression de travail (Ps) - à l'encontre de la force (F1) d'un ressort de distributeur (28) - ; et
- à l'encontre de l'action de la première surface de piston de commande (A1) - une deuxième surface de piston de commande (A2), qui est chargée par la pression de commande (Pc), caractérisée
en ce que le dispositif de rappel (10) comprend un ou plusieurs ressorts (15, 16) pour produire une force de rappel totale progressivement croissante, en présence d'une variation dirigée vers de plus petits volumes de refoulement (Q), et en
ce que la deuxième surface du piston de commande (A2) est plus grande que la première surface du piston de commande (A1).
2. Pompe à régulation de puissance selon la revendication 1, caractérisée en ce que les ressorts (15, 16) entrent en action l'un après l'autre au cours de leur débattement élastique.
3. Pompe à régulation de puissance selon la revendication 2, caractérisée en ce que le rapport des grandeurs de la première et de la deuxième surface du piston de commande (A1 : A2) ainsi que les caractéristiques élastiques (K1, K2) et que les courses de débattement des ressorts (15, 16) du dispositif de rappel (10) sont choisis de manière à obtenir une courbe de limitation de puissance constante qui se rapproche d'une hyperbole dans le diagramme QP.
4. Pompe à régulation de puissance selon une des revendications 1 à 3, caractérisée en ce que l'organe de commande (22) du distributeur (20) présente deux arêtes de commande sur une portée de piston (24), qui coopèrent avec un perçage de commande (25) et qui, dans la position moyenne de la portée de piston (24), laissent un débit de fuite étranglé s'écouler d'une conduite de pression (19) vers une conduite d'écoulement (14).
EP19850109781 1985-08-03 1985-08-03 Réglage de puissance ou limitation de couple Expired - Lifetime EP0211096B1 (fr)

Priority Applications (2)

Application Number Priority Date Filing Date Title
EP19850109781 EP0211096B1 (fr) 1985-08-03 1985-08-03 Réglage de puissance ou limitation de couple
DE8585109781T DE3578196D1 (de) 1985-08-03 1985-08-03 Leistungsregelung oder drehmomentbegrenzung.

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP19850109781 EP0211096B1 (fr) 1985-08-03 1985-08-03 Réglage de puissance ou limitation de couple

Publications (2)

Publication Number Publication Date
EP0211096A1 EP0211096A1 (fr) 1987-02-25
EP0211096B1 true EP0211096B1 (fr) 1990-06-13

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Family Applications (1)

Application Number Title Priority Date Filing Date
EP19850109781 Expired - Lifetime EP0211096B1 (fr) 1985-08-03 1985-08-03 Réglage de puissance ou limitation de couple

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EP (1) EP0211096B1 (fr)
DE (1) DE3578196D1 (fr)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102021201409A1 (de) 2021-02-15 2022-08-18 Robert Bosch Gesellschaft mit beschränkter Haftung Verdrängermaschine mit einer Messvorrichtung für das Verdrängungsvolumen

Families Citing this family (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE29922559U1 (de) 1999-12-22 2000-03-02 Thüringer Bauholding GmbH, 63654 Büdingen Zeltgerüst
US10570893B2 (en) * 2015-05-29 2020-02-25 Kanzaki Kokyukoki Mfg. Co., Ltd. Hydraulic pump and detachable servo unit
CN117189533A (zh) * 2023-08-18 2023-12-08 江苏恒立液压科技有限公司 一种柱塞泵

Family Cites Families (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2921560A (en) * 1957-09-23 1960-01-19 New York Air Brake Co Engine control
US3093081A (en) * 1959-01-29 1963-06-11 New York Air Brake Co Pumping device
SE345501B (fr) * 1967-01-16 1972-05-29 Wytwornia Sprzetu Komunikacyjn

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102021201409A1 (de) 2021-02-15 2022-08-18 Robert Bosch Gesellschaft mit beschränkter Haftung Verdrängermaschine mit einer Messvorrichtung für das Verdrängungsvolumen

Also Published As

Publication number Publication date
EP0211096A1 (fr) 1987-02-25
DE3578196D1 (de) 1990-07-19

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