EP0179249B1 - Hydraulische Steuerung - Google Patents

Hydraulische Steuerung Download PDF

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Publication number
EP0179249B1
EP0179249B1 EP85111444A EP85111444A EP0179249B1 EP 0179249 B1 EP0179249 B1 EP 0179249B1 EP 85111444 A EP85111444 A EP 85111444A EP 85111444 A EP85111444 A EP 85111444A EP 0179249 B1 EP0179249 B1 EP 0179249B1
Authority
EP
European Patent Office
Prior art keywords
pressure
valve
control
supply
hydraulic 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 - Lifetime
Application number
EP85111444A
Other languages
German (de)
English (en)
French (fr)
Other versions
EP0179249A2 (de
EP0179249A3 (en
Inventor
Hubert Häussler
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.)
Beringer-Hydraulik AG
Original Assignee
Beringer-Hydraulik AG
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Beringer-Hydraulik AG filed Critical Beringer-Hydraulik AG
Priority to AT85111444T priority Critical patent/ATE51089T1/de
Publication of EP0179249A2 publication Critical patent/EP0179249A2/de
Publication of EP0179249A3 publication Critical patent/EP0179249A3/de
Application granted granted Critical
Publication of EP0179249B1 publication Critical patent/EP0179249B1/de
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

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Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B66—HOISTING; LIFTING; HAULING
    • B66B—ELEVATORS; ESCALATORS OR MOVING WALKWAYS
    • B66B1/00—Control systems of elevators in general
    • B66B1/24—Control systems with regulation, i.e. with retroactive action, for influencing travelling speed, acceleration, or deceleration

Definitions

  • the invention relates to a hydraulic control according to the preamble of claim 1.
  • This hydraulic control has a metering valve arranged in the inlet to the consumer with an adjustable throttle and a pressure compensator, the balance piston for measuring the pressure drop at the adjustable throttle and for generating a pressure-dependent one hydraulic control pressure on one side (inlet pressure side) with the pressure in front of the throttle (inlet pressure) and on the other side (load pressure side) with the pressure behind the throttle (load pressure).
  • a flow control valve (control valve) is arranged in the inlet between the pump and metering valve, which is hydraulically pilot-controlled depending on the control pressure of the pressure compensator.
  • control pressure is produced in that the control pressure chamber of the control valve is supplied with oil on the inlet pressure side via a throttle and is connected to the tank via the control edge of the pressure compensator. This creates a permanent oil flow through the control room via the pressure compensator to the tank even in static operation of the hydraulic control.
  • the known hydraulic control is therefore lossy.
  • the hydraulic control further have the features of the characterizing part in claim 1 and thus the pressure compensator receives a double-edge control via which the control chamber of the control valve can be connected on the one hand to a reference pressure line and on the other hand to the tank line.
  • the oil flow is only for the hydraulic adjustment of the control valve. This oil flow is irrelevant in terms of its power loss and oil consumption. However, as long as the control valve is not adjusted, no oil flows into or out of the control room.
  • the control valve is preferably designed as a bypass valve.
  • the valve connects the inlet to the tank channel. Its control piston is acted on one side by the inlet pressure and a spring and on the other side by the control pressure of the pressure compensator.
  • This valve has the advantage in this application that it is suitable for both upward operation and for downward operation without further changeover.
  • the pressure compensator is connected to a standard pressure line.
  • the set pressure in this set pressure line is converted into a control pressure depending on the pressure difference existing on the balance piston.
  • the default pressure line can either be connected to the inlet between the pump and the adjustable throttle or to the load pressure side of the throttle.
  • the specified pressure line of the pressure compensator is connected via a shuttle valve on the one hand to the inlet in front of the adjustable throttle and on the other hand to the consumer channel behind the adjustable throttle. This ensures that the higher pressure is always available to control the control valve. This is particularly important if the inlet between the control valve and the adjustable throttle is essentially depressurized at a standstill or during slow lowering operation. In this case, the low pressure would no longer be sufficient to actuate the control valve.
  • the invention further provides that the balance piston of the pressure compensator is clamped with an adjustable spring force such that the zero position of the balance piston can be adjusted.
  • the zero position is the position of the balance piston that the balance piston assumes when none of its sides is pressurized with oil pressure.
  • the spring force is adjustable in such a way that the balance piston connects the control connection of the flow control valve with the default pressure line in one zero position and covers the control connection in the other zero position or connects it slightly to the tank connection.
  • control valve In the case of upward operation, the control valve is pressurized with the default pressure in the sense of closing and building up a pressure in the inlet until the inlet pressure overcomes the spring preload and the load pressure on the balance piston and closes the control line from the default pressure line and possibly from the tank line opens.
  • the load pressure initially outweighs the pressure in the supply line. Therefore, the load pressure shifts the balance piston against the biasing force of the now switched spring clamping, and it becomes the control pressure chamber is connected to the default pressure line in the sense of closing the flow control valve and increasing the pressure.
  • this pressure With increasing pressure in the supply line, this pressure now acts in the same sense as the spring clamping on the balance piston against the load pressure in such a way that the balance piston closes the control line with respect to the specified pressure line and possibly opens with respect to the tank line, so that the flow control valve in the sense of opening to the tank and the lowering of the pressure in the supply line is actuated.
  • a stop can be provided which has the function of restricting the movement of the balance piston and, in particular, only allowing a throttled opening between the specification pressure line and the control line. This can dampen the movement of the control valve.
  • the adjustability of the spring clamping of the balance piston means that the hydraulic control according to this invention with the same components is effective both in upward and downward operation.
  • the spring forces are preferably set as a function of the elevator control, and preferably in that at least one of the counter-bearings of the clamping spring can be adjusted between two positions by means of a suitable force transmitter.
  • the counter bearing is preferably designed as a hydraulically actuated piston which can be hydraulically adjusted by the elevator control depending on the direction of travel of the consumer.
  • the metering valve is also hydraulically pilot-controlled, in that it is acted upon on the one hand by the inlet pressure upstream of the metering valve and on the other hand by a controllable counter pressure.
  • the metering piston is designed as a differential piston, the small piston piece of which is supplied with the inlet pressure on the front side and has a collar acting as a seat valve, through which the load channel is leak-free compared to the inlet channel can be sealed.
  • the annular surface between the small and the large piston part is acted upon by the load pressure.
  • the large piston area is acted upon by the controllable back pressure.
  • the back pressure is derived from the load pressure via the previously mentioned shuttle valve.
  • the thin end of the throttle piston has an annular groove directly in front of its seat, which is connected to the load pressure detection channel and the shuttle valve by a load pressure detection channel in the metering piston and an annular groove in the thick piston piece.
  • This configuration of the metering piston ensures that the load pressure is applied on the one hand to the counterpressure side of the metering piston in the closing direction and on the other hand to the balance piston of the pressure compensator in the closing direction of the flow control valve before the metering valve has connected the inlet and the consumer line.
  • a pressure corresponding to the load pressure can build up in the feed line before the connection between the load line and the feed line is established.
  • the back pressure chamber of the metering valve is - as already mentioned - connected to the load pressure signaling channel via an inlet throttle. Furthermore, the back pressure chamber is connected to the tank via an outlet throttle and a check valve.
  • the metering piston can be actuated hydraulically by opening the shut-off valve and the ratio of the inlet throttle to the outlet throttle.
  • the special feature of the hydraulic control for elevators according to the invention is that the pressure compensator can be used to set a specific pressure difference between the consumer line and the inlet.
  • the driving behavior of the consumer essentially depends on the stroke movement of the metering piston. This stroke movement is predetermined by the inlet throttle and the outlet throttle, so that a load-independent driving behavior and constant accelerations and decelerations can be expected.
  • the consumer 1 is shown schematically.
  • the elevator cylinder 3 with the piston 2 is shown.
  • the hydraulic pump 4 is driven by the motor 5.
  • the hydraulic oil is removed from the tank 6 and pumped into the line 7, which is referred to below as the inlet.
  • the metering valve 8 with a hydraulically controlled throttle is located in the inlet 7.
  • the control valve 9, also called the flow control valve, with control pistons 10, 11 controls the pressure build-up in the inlet 7.
  • a bypass to the tank channel 13 is opened or closed.
  • the thin piston piece 11 has control grooves which establish the connection between the inlet 7 and the tank connection 13.
  • the thin piston end is loaded by spring 12 and by the inlet pressure.
  • the thick piston piece 10 is acted upon by the control pressure in the control chamber 14. It can - what here is not shown - a switchable pressure relief of the control chamber 14 may be provided. This pressure relief takes place in particular when a star-delta switchable motor 5 is used, which starts when the star connection is made.
  • a check valve 16 is also provided in the inlet 7. The check valve closes when the pump 4 is at a standstill, that is to say when the elevator is at a standstill and in the lowering mode.
  • a throttle valve is provided as the metering valve 8 and is hydraulically pilot-controlled. Further details on this will be described later with reference to FIG. 2.
  • the pressure in the control room 14 is controlled via the control line 15 by the pressure compensator 17.
  • the pressure compensator 17 has a weighing piston (18) which is clamped between the springs 22 and 23.
  • the counter bearing of the spring 23 is formed by a differential piston 24, 25.
  • the differential piston can be pressurized via adjusting line 27 and adjusting valve 30 in adjusting pressure chamber 28.
  • the opposite space of the cylinder is connected to tank 6 via tank connection 13.
  • the pressure relief of the adjusting pressure chamber 28 takes place via the narrow throttle 26 in the piston 25.
  • the one end position of the piston 24, 25 can be fixed by means of the adjusting screw 29.
  • the end pistons of the balance piston 18 form hydraulic control chambers, of which the load pressure chamber 20 is acted upon by the load pressure or consumer pressure via the load pressure signaling channel 34 and the inlet pressure is supplied to the inlet pressure chamber 19 via the inlet pressure signaling channel 33.
  • the balance piston carries out a control movement.
  • the central control collar 35 of the pressure compensator 17 interacts with this control movement with the outlet of the control line 15 and controls with its two control edges the control line 15 with respect to the standard pressure connection 21 and the tank connection 13.
  • the standard pressure line 31 is connected to a shuttle valve 32.
  • the shuttle valve 32 is located on the one hand on the inlet pressure reporting channel 33 and on the other hand on the load pressure reporting channel 34. The higher of these pressures is applied via line 31 and port 21 of the pressure compensator as the default pressure.
  • the metering valve 8 has the metering piston 36 designed as a differential piston.
  • the thinner end 37 of the metering piston has the control notches 38, via which the connection between the inlet 7 and the consumer channel 20 is established.
  • the thin end 37 has a collar 39 which forms a seat valve with the valve seat 40.
  • the metering piston 36 can seal the consumer channel 20 against the inlet channel 7 without leakage, which is particularly important when the machine is at a standstill in order to prevent the consumer from dropping unintentionally, i.e. to avoid the elevator car.
  • the thin end 37 has the annular channel 41 immediately following the collar 39.
  • annular groove 43 in the thicker end 44 of the metering piston 36 This is connected via signal line 42 to an annular groove 43 in the thicker end 44 of the metering piston 36.
  • the annular groove is enclosed by suitable dynamic seals and is connected to the load pressure signaling channel 34, which on the one hand leads into the shuttle valve already described and on the other hand - not shown in FIG. 2 - into the pressure compensator 17 and to the adjusting valve 30.
  • the special arrangement of the annular groove 41 in the metering piston 36 has the effect that, immediately after the metering piston is lifted from the seat 40, the load pressure in the consumer channel 20 is reported to the pressure compensator via the channel system 42, 34, even before the consumer channel 20 via notches 38 with the inlet 7 connected is.
  • the metering piston is loaded on its large piston surface by the spring 45. Furthermore, the metering piston has a connecting channel 46 with throttle 47, which loads the control pressure side 48 of the metering valve 8 with the consumer pressure, even if the metering valve on the seat 40 is closed leak-free. This ensures that, when the pump is at a standstill, the metering piston is pressed onto its seat without leakage.
  • the large piston side is connected to the pressure converter 52, which is shown schematically in FIG. 1 and in FIG. 2 by dashed lines.
  • the pressure converter 52 consists of an inflow plug 49 and an outlet throttle 50, a check valve 55 and a seat valve 51, through which the tank line 13, in which the outlet throttle 50 is located, can be opened and closed without leakage.
  • the control line 53 of the metering valve is connected via the supply throttle 49 on the one hand to the preset pressure line 31 of the shuttle valve 32 and on the other hand via the seat valve 51 and the discharge throttle 50 to the tank.
  • the feed throttle 49 and the outlet throttle 50 can be adjusted to a constant oil flow. They are therefore preferably designed as adjustable flow control valves. After the flow ratio has been set, the control pressure in the control chamber 48 is only dependent on the set pressure. It should be noted that the throttle 47 in the metering piston 44 is very small compared to the feed throttle 49.
  • the motor 5 and the pump 4 are at a standstill.
  • the consumer exerts a pressure in the consumer channel 20.
  • This pressure is applied to the large piston side of the metering piston 44, ie control chamber 48, via the load channel 46 and the throttle point 47.
  • the metering piston 44 seals the consumer channel 20 on its collar 39 from the supply channel 7 without leakage.
  • the inlet 7 is essentially depressurized.
  • the ring channel 41 on the metering piston 44 is also depressurized.
  • the check valve 55 in the pressure converter 52 prevents the oil from flowing back from the control chamber 48 via line 53 into the control. Therefore, the control chamber 14 of the control valve 9 is depressurized.
  • the Piston 10, 11 thus opens the supply channel 7 due to the spring force 12 to the tank connection 13.
  • the motor 5 and the pump 4 are started.
  • the valve 51 is switched.
  • the valve 30 remains in the switch position shown.
  • Other switching options for controlling the acceleration and for controlling the creep speed when entering the target are not taken into account.
  • the possibility of depressurizing the control chamber 14 when the motor 5 starts in star connection is also not taken into account.
  • the control chamber 14 Since the load pressure signaling channel 34 and inlet pressure signaling channel 33 are initially still depressurized, the control chamber 14 is also depressurized.
  • the spring 12 presses the piston 10 against the stop screw 54. This is set such that the oil flow is throttled at the control grooves and a pressure of approximately 3 to 6 bar is established in the inlet 7.
  • This inlet pressure is applied to the balance piston 18 on the inlet pressure side 19 via the inlet pressure signaling channel 33.
  • This feed pressure also reaches the pressure converter 52 via the change-over valve 32 and the predetermined pressure line 31 and, via this and the control line 53, to the control side 48 of the metering piston 44.
  • the inlet throttle 49 and the outlet throttle 50 are preferably designed as flow regulators and are set such that the oil flow via the inlet throttle 49 is about half the size of the oil flow through the outlet throttle 50. This causes the metering piston 44 to be relieved of pressure on its control pressure side 48 and moved to the right by the inlet pressure in inlet 7. In doing so, it displaces the oil volume in the control chamber 48 via the outlet throttle 50.
  • the annular space 41 is connected to the consumer channel 20.
  • the consumer pressure is therefore given via the annular space 41 and the signal line 42 into the annular groove 43 and from here via the load pressure signaling channel 34 on the one hand to the shuttle valve 32 and on the other hand to the load pressure side 20 of the pressure compensator 17.
  • the counter bearing piston 25 lies on the adjusting screw 29.
  • the springs 22 and 23 are dimensioned such that the spring force 22 predominates in this position and acts on the balance piston in the direction of the stop 57. Since at the same time the load pressure side 20 of the pressure compensator is pressurized in the direction of the spring force 22, the balance piston rests on the stop 57. As a result, the preset pressure connection 21 is opened towards the control connection 15 and the control chamber 14 of the control valve 9 is acted upon with the preset pressure. The set pressure is selected by the shuttle valve 32 between the inlet pressure and the consumer pressure, whichever is the higher.
  • control piston 10 of the control valve - in FIG. 1 - moves to the left in the sense that the supply channel 7 is closed off from the tank channel 13.
  • the pressure in the inlet 7 continues to build up. Since this inlet pressure is also applied to the pressure compensating piston on the inlet pressure side 19 via the inlet pressure signaling channel 33, this inlet pressure counteracts the spring force 22 and the load pressure on the load pressure side 20 in the sense that the control line 15 is initially closed with respect to the preset pressure connection 21 and then with the tank connection 13 is connected. If the pressure gradient between the load pressure chamber 20 and the inlet pressure chamber 19 and thus also the pressure gradient between the consumer duct 20 and the inlet 7 becomes too large, i.e.
  • the upward travel is ended by switching off the solenoid valve 51.
  • the devices and circuits for bringing about a creep speed before reaching the destination are not described and illustrated.
  • the motor 5 is also switched off with a certain delay. All elements thus assume the rest position shown in the drawing.
  • the consumer channel 20 is blocked off from the inlet 7 again leak-free by collar 39 on seat 40.
  • the consumer pressure builds up again in the control room 48.
  • the check valve 55 prevents the oil from flowing back from the control chamber 48 into the pilot control area.
  • the pressure drop from the inlet channel 7 to the consumer 20 can be predetermined by the adjusting screw 29 on the counter bearing.
  • the flow at the metering valve can be changed almost in a ratio of 1: 2. This means that one and the same metering valve version can be used for a wide range of applications.
  • valves 30 and 51 are simultaneously switched to flow.
  • the motor 5 with pump 4 remains out of operation.
  • Check valve 16, through which the pump is connected to the inlet 7, is closed in the direction of the pump by the acting spring.
  • the control chamber 48 is relieved of pressure by switching the valve 51.
  • the metering piston 44 therefore moves under the pressure of the consumer, which acts on the mutual annular surface of the metering piston 44 - in FIG. 2 - to the right.
  • the metering piston 44 lifts off from the seat 40 with its collar 39.
  • the load pressure in the consumer channel 20 via ring channel 41, avoidance channel 42, ring groove 43 and load pressure reporting line 34, shuttle valve 32, line 31, on the one hand, via pressure converter 52 to the control side 48 of the metering piston and, on the other hand, via connection 21 as the default pressure to the pressure compensator and furthermore via line 34 given on the one hand to the adjusting valve 30 and on the other hand to the load pressure side 20 of the pressure compensator.
  • the pressure ratio at the metering valve for the lowering travel can be set independently of that during the upward travel.
  • the pressure ratio for the upward travel is set by adjusting screw 29.
  • the pressure balance piston now acts - in FIG. 1 - to the left the inlet pressure on the inlet pressure side 19 and the spring force, which was reversed in its direction of action by shifting the counter bearing 24, and to the right the load pressure on the load pressure side 20.
  • the load pressure is at lowering higher than the inlet pressure. Therefore, the balance piston 18 - in FIG. 1 - is shifted to the right as long as the metering piston 44 is still closed.
  • the set pressure (line 31) thus reaches the control line 15 and the control chamber 14 of the control valve via the pressure compensator. This closes the bypass from inlet 7 to the tank line 13 against the spring force 12.

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  • Engineering & Computer Science (AREA)
  • Automation & Control Theory (AREA)
  • Fluid-Pressure Circuits (AREA)
  • Valve Device For Special Equipments (AREA)
  • Types And Forms Of Lifts (AREA)
EP85111444A 1984-09-15 1985-09-10 Hydraulische Steuerung Expired - Lifetime EP0179249B1 (de)

Priority Applications (1)

Application Number Priority Date Filing Date Title
AT85111444T ATE51089T1 (de) 1984-09-15 1985-09-10 Hydraulische steuerung.

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE19843434014 DE3434014A1 (de) 1984-09-15 1984-09-15 Hydraulische steuerung
DE3434014 1984-09-15

Publications (3)

Publication Number Publication Date
EP0179249A2 EP0179249A2 (de) 1986-04-30
EP0179249A3 EP0179249A3 (en) 1987-09-30
EP0179249B1 true EP0179249B1 (de) 1990-03-14

Family

ID=6245550

Family Applications (1)

Application Number Title Priority Date Filing Date
EP85111444A Expired - Lifetime EP0179249B1 (de) 1984-09-15 1985-09-10 Hydraulische Steuerung

Country Status (7)

Country Link
US (1) US4676140A (ja)
EP (1) EP0179249B1 (ja)
JP (1) JPH0615881B2 (ja)
AT (1) ATE51089T1 (ja)
CA (1) CA1243585A (ja)
DE (2) DE3434014A1 (ja)
DK (1) DK167863B1 (ja)

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DE3536219A1 (de) * 1985-10-10 1987-04-16 Heilmeier & Weinlein Hydraulische steuervorrichtung
FI83204C (fi) * 1987-11-04 1991-06-10 Kone Oy Foerfarande och anordning foer foerbaettring av verkningsgraden hos en motorstyrd hydraulhiss.
US5082091A (en) * 1990-01-19 1992-01-21 Otis Elevator Company Hydraulic elevator control
US5014824A (en) * 1990-01-19 1991-05-14 Otis Elevator Company Hydraulic elevator control valve
US5212951A (en) * 1991-05-16 1993-05-25 Otis Elevator Company Hydraulic elevator control valve
IT1248792B (it) * 1991-05-20 1995-01-30 Gmv Martini Spa Circuito idraulico per ascensori, montacarichi e simili, con sicurezza intrinseca
DE4219552C2 (de) * 1992-06-15 1996-03-07 Rexroth Mannesmann Gmbh Vorgesteuertes Zwei-Wegeventil mit einstellbarer, druckunabhängiger Schließzeit
DE4223389C2 (de) * 1992-07-16 2001-01-04 Mannesmann Rexroth Ag Steueranordnung für mindestens einen hydraulischen Verbraucher
US5289901A (en) * 1992-08-03 1994-03-01 Otis Elevator Company Hydraulic elevator pressure relief valve
JP3175418B2 (ja) * 1993-08-18 2001-06-11 三菱電機株式会社 油圧エレベーターの制御装置
US5374794A (en) * 1993-12-09 1994-12-20 United States Elevator Corp. Elevator control valve assembly
CZ286074B6 (cs) * 1996-01-30 2000-01-12 Mannesmann Rexroth Ag Hydraulické zařízení k ovládání toku tlakového prostředku
US5992573A (en) * 1997-09-24 1999-11-30 Blain; Roy W. Elevator up start
US6276135B1 (en) * 1999-04-29 2001-08-21 Argus Machine Co. Ltd. Self-contained hydraulic ESD system
US7209806B2 (en) * 2003-07-25 2007-04-24 Timm Miguel A Self-contained electronic pressure monitoring and shutdown device
US9581267B2 (en) 2011-04-06 2017-02-28 David John Kusko Hydroelectric control valve for remote locations
DE102017114704B4 (de) * 2017-06-30 2019-10-10 Vemcon Gmbh Modulare Hydraulikeinheit, Fahrzeug und Steuersystem für eine Arbeitsmaschine
EP3444213A1 (de) * 2017-08-17 2019-02-20 Blain Hydraulics GmbH Hydraulischer aufzug

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Also Published As

Publication number Publication date
DK167863B1 (da) 1993-12-27
CA1243585A (en) 1988-10-25
JPS61112801A (ja) 1986-05-30
DE3434014A1 (de) 1986-03-20
DK410485D0 (da) 1985-09-10
EP0179249A2 (de) 1986-04-30
EP0179249A3 (en) 1987-09-30
DE3576584D1 (de) 1990-04-19
ATE51089T1 (de) 1990-03-15
JPH0615881B2 (ja) 1994-03-02
US4676140A (en) 1987-06-30
DK410485A (da) 1986-03-16

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