EP1092095A1

EP1092095A1 - Hydraulic circuit

Info

Publication number: EP1092095A1
Application number: EP99936360A
Authority: EP
Inventors: Thomas Weickert; Erich Adlon
Original assignee: Mannesmann Rexroth AG; Bosch Rexroth AG
Current assignee: Bosch Rexroth AG
Priority date: 1998-06-29
Filing date: 1999-05-31
Publication date: 2001-04-18
Anticipated expiration: 2019-05-31
Also published as: WO2000000747A1; JP4520041B2; EP1092095B2; DE59904746D1; EP1092095B1; JP2002519596A; DE19828963A1; US6367365B1; KR20010071687A; KR100636863B1

Abstract

The invention relates to a load-independent flow regulation circuit for controlling at least one low load consumer (6) and one high load consumer (4), wherein each consumer (4, 6) is allocated a metering orifice (14a, 14b) and a downstream pressure regulator (16a, 16b) to constantly maintain pressure drop above the metering orifice (16a). A controllable bypass duct (32) is allocated to the pressure regulator of the low load consumer through which the pressure regulator of said consumer may be circumvented.

Description

Description Hydraulic circuit

The invention relates to a hydraulic circuit for controlling at least one load-lower and one load-higher consumer according to the preamble of patent claim 1.

Such circuits (also called load-sensing circuits) are used, among other things, to control mobile machines, for example excavators. Hydraulically operated units of the working machine, for example a rotating mechanism, the drive, a bucket, a stick or a clamping device mounted on the excavator boom, are controlled via the central circuit.

Such a load-sensing circuit is known, for example, from EP 0 566 449 AS. This circuit has a variable displacement pump which can be regulated in such a way that it produces a pressure at its output which is a certain difference above the maximum load pressure of the hydraulic consumers. For control purposes, a load-sensing controller is provided, which can be acted upon by the pump pressure in the direction of reducing the stroke volume and by the highest pressure at the consumers, and by a compression spring in the direction of increasing the stroke volume. The difference between the pump pressure and the highest load pressure that occurs with the variable displacement pump corresponds to the force of the aforementioned compression spring.

Each of the consumers is assigned an adjustable orifice plate with a downstream pressure balance, via which the pressure drop at the orifice plate is kept constant is so that the amount of hydraulic fluid flowing to the respective consumer depends on the opening cross section of the measuring orifice and not on the load pressure of the consumer or on the pump pressure. In the case in which the variable displacement pump delivers at maximum volume and the hydraulic fluid flow is still not sufficient to maintain the specified pressure drop across the orifice plates, the pressure compensators of all actuated hydraulic consumers are adjusted in the closing direction so that all hydraulic fluid flows are closed the individual consumers by the same proportion. This means that with a pressure compensator connected downstream, the volume flows to the consumers are always in the ratio of the opening cross-sections of the orifice plates. Because of this load-independent flow distribution (LÜDV), all controlled consumers move at a speed that is reduced by the same percentage.

The variable displacement pump mentioned at the outset is usually equipped with a pressure control and a power control, by means of which the maximum possible pump pressure or the maximum output (excavator output) that can be output by the variable displacement pump can be set. These pressure and power controls are superimposed on the load sensing control.

With a control arrangement of the type described above, problems can arise if a hydraulic consumer works against a practically infinite resistance. This can be the case, for example, if the hydraulic consumer is a bucket that is moved to a stop. When driving to the limit, a pressure builds up on the corresponding hydraulic consumer, which corresponds approximately to the maximum pressure (excavator output) specified by the pressure control. Now becomes another hydraulic consumer, If, for example, a travel drive or a boom is actuated, it can only be moved at a lower speed because, due to the high pressure at the first-mentioned consumer (bucket), the power control of the variable displacement pump responds to the other hydraulic consumer (travel drive) even with low hydraulic fluid flows.

To overcome this disadvantage, a control arrangement is disclosed in the applicant's W095 / 32364, by means of which, when a limit load pressure is exceeded, only the load pressure of the lower-load hydraulic consumer is reported to the load-sensing controller of the variable pump. This limit load pressure is selected so that the supply of the other hydraulic consumer is guaranteed. In the case of W095 / 32364, this is achieved in that the spring chamber of the pressure compensator of the lower-load consumer can be connected to the tank via a pressure relief valve arrangement. If a limit load pressure is exceeded, the pressure relief valve opens the connection to the tank, so that the spring chamber of the pressure compensator of the load-lower consumer is relieved and the control piston is brought into its open position, in which the load pressure of this consumer is reported to the load pressure reporting line.

A disadvantage of this control arrangement is that a partial volume flow is discharged to the tank and therefore cannot be used for consumer control. The efficiency of this control is therefore comparatively low. Another disadvantage is that the return of the hydraulic fluid to the tank generates heat in the system and thus pump performance is destroyed. In contrast, the invention has for its object to provide a control arrangement through which an adequate supply of all consumers is guaranteed with minimal expenditure on device technology.

This object is achieved by a hydraulic circuit with the features of claim 1.

The measure of providing a bypass channel through which the pressure compensator downstream of the measuring orifice can be bypassed means that it is not necessary to remove the pressure compensator in order to limit the system pressure or to drain hydraulic fluid into the tank. The resulting system pressure can be predetermined by selecting the bypass cross-section accordingly. Due to the reduced system pressure, the load-lower consumer can be supplied with a larger amount of hydraulic fluid, which can be implemented, for example, in a boom speed increase or the like.

A circuit of particularly simple construction is obtained if the measuring orifice upstream of the pressure compensator is formed by a proportional directional control valve, the bypass channel being controllable as a function of the valve spool position of the proportional directional control valve. Because the bypass channel is activated depending on the control of the proportional valve, the individual pressure compensator only works in the fine control range, in which comparatively low hydraulic fluid volume flows flow through the pressure compensator.

The structure can be further simplified if the bypass channel is formed in the valve slide of the proportional directional control valve and can be opened by a control edge of the valve slide bore. In order to prevent backflow from the consumer through the bypass duct, a check valve arrangement is provided in this.

In a preferred variant of the invention, two work connections of a consumer are controlled via the proportional valve. In some cases, for example in the case of double-acting hydraulic cylinders, it is sufficient if the bypass channel is assigned to only one of the work connections, so that, for example, the bypass is flowed through in the lifting function. Of course, it is also possible to assign bypass channels to both work connections.

As already mentioned above, it can be advantageous if the bypass channel is only opened after a certain stroke of the proportional valve, so that no bypass flow occurs at the start of the regulation.

The valve spool of the proportional directional control valve is preferably formed with a central speed part and two external directional parts, each of which is assigned to a connection of the consumer. The bypass channel extends within the valve spool from the speed section to the direction section, so that the pressure compensator is bypassed.

The pressure loss in the bypass channel can be minimized if it ends with oblique and radial bores in the outer circumference of the valve spool.

Other advantageous developments of the invention are the subject of the further subclaims. Preferred exemplary embodiments of the invention are explained in more detail below with the aid of schematic drawings. Show it:

Figure 1 is a circuit diagram of a circuit according to the invention with a bypass channel.

FIG. 2 shows a valve disk of a valve block for a circuit according to FIG. 1;

3 shows a section through a valve segment for a circuit according to FIG. 1;

Fig. 4 is a detailed view of the valve segment of Fig. 3 and

Fig. 5 is a diagram to illustrate the system pressure build-up when driving a load higher and a load lower consumer.

1 shows part of a circuit diagram for a hydraulic circuit for controlling a mobile implement, for example an excavator. This excavator has several consumers, such as a boom, a bucket, a stick, a chassis drive and a slewing gear drive, which are supplied with hydraulic fluid by a variable displacement pump 2. In the embodiment shown in FIG. 1, a cylinder 4 for actuating a bucket and a cylinder 6 for actuating the excavator boom are shown schematically as consumers.

The stroke volume of the variable displacement pump is set by means of a load-sensing controller 8 which, depending on the pump pressure on the one hand and the highest load pressure at the consumers 4, 6 and the force of a pressure spring 10 on the other hand controls the displacement of the variable displacement pump. The hydraulic fluid delivered by the variable displacement pump is led to the two consumers 4 and 6 via a pump line 12 with branch lines 12a, 12b.

An adjustable orifice 14a, 14b is formed in each branch of the pump line 12 (12a, 12b). As will be explained in more detail below, these orifices are

14a, 14b designed as speed parts of a proportional valve.

A pressure compensator 16a, 16b is connected downstream of each measuring orifice 14a, 14b. The control piston of these 2-way pressure compensators is pressurized in the opening direction via a control line 18 with the pressure downstream of the measuring orifice 14a, 14b and in the closing direction via a load control line 20 with the highest load pressure, which is tapped from a load pressure signaling line 22. The highest load pressure is also fed to the load-sensing controller 8 via this.

A working line 24a, 24b leads to the respective consumers 4 and 6 from the outlet connection of the pressure compensator 16a, 16b. The load pressure of the consumers 4, 6 is tapped via branch lines 26a, 26b and led to a shuttle valve 28, to the output of which the load pressure reporting line 22 is connected is.

The adjustable measuring diaphragms 14a, 14b are controlled via manually operable control devices 30a, 30b which are operatively connected to the measuring diaphragms 14a and 14b.

A circuit of the type described above realizes a classic "LUDV" circuit, in which the pressure drop across the measuring orifices 14a, 14b is kept constant regardless of the load pressure. When the full pump capacity is exhausted, both pressure compensators 16a, 16b are usually reduced, so that the hydraulic fluid volume flow to the two consumers 4, 6 is reduced by the same percentage. As already described at the beginning, a problem can occur with these circuits when the higher load consumer (bucket 4) is moved to the stop, so that the load pressure of this consumer is in the range of the maximum pump pressure. If you now also switch on a load-lower consumer, the volume flow of the load-lower consumer goes back to a value that is predetermined by the maximum pump output. A large part of the performance is destroyed in the regulating pressure compensator of this consumer.

In order to prevent this, in the control system shown in FIG. 1 the bypass load 32 is assigned to the lower load b, which enables the pressure compensator 16a to be bypassed. The bypass duct 32 branches off downstream of the measuring orifice 14a and opens into the working line 24a to the consumer 6. A suitable control device 34 is provided in the bypass duct 32, which blocks the bypass duct 32 in the basic position and opens depending on the opening cross section of the measuring orifice 14a. With this circuit, the hydraulic fluid volume flow to the consumer 6 is not regulated by the pressure compensator 16a, so that a lower system pressure than in a system without a bypass channel 32 is established. This makes it possible to extend the boom 6 at a higher speed. The switching device provided with the reference numeral 34 can be any device that is suitable for shutting off the bypass channel 32 and opening it as a function of the control of the measuring orifice 14a. FIG. 2 shows the circuit diagram of a valve disk 35 of a valve block for realizing the circuit shown in FIG. 1. The valve disc 35 contains the pressure compensator 16a, a proportional valve 36, by means of the speed part of which the measuring orifice 14a is formed and the bypass channel 32, as well as the other connecting lines of the hydraulic elements described in more detail below. In the exemplary embodiment shown in FIG. 2, in addition to the measuring orifice 14a, a directional part for controlling the consumers A, B and for controlling the bypass channel 32 are also integrated in the proportional valve 36.

The proportional valve 36 has a pump connection P, two working connections A, B, which are connected to the cylinder chambers of a differential cylinder b or to a hydraulic motor. Furthermore, an output port P1 to the pressure compensator 16a, a bypass port U, two input ports R, S of the directional part and a tank port T are formed on the proportional valve 36.

The two end faces of the valve slide 38 of the proportional valve 36 are prestressed into their basic position by two compression springs 41a, 41b. In this basic position, the ports P, A, B, U and S are closed, while the ports Pl and R are connected to the tank.

The end faces of the valve _spool 38 are acted upon by control pressures P _sτ , so that it can be moved out of its spring-loaded basic position.

The output port P1 is connected to the input port Q of the pressure compensator 16a via the pump line 12a. As already explained above, branches of the Pump line 12a from the control line 18, via which the pressure downstream of the measuring orifice 14a (proportional valve 36) is reported to the left end face of the pressure compensator 16a in FIG. 2. The load pressure of the consumer 6 is connected via the load signaling line 20 to the load pressure signaling line 22 and led to the spring side of the pressure compensator 16a. The output connection C of the pressure compensator 16a is connected via lines 40, 42 to the input connections R and S of the directional part. In the lines 40, 42 there are two non-return valves 56a, 56b which prevent the hydraulic fluid from flowing back from the directional part to the pressure compensator 16a.

The tank connection T is connected to the tank via a tank line 44. By means of the pressure compensator 16a, when the proportional valve 36 is activated, the pressure drop across the measuring orifice 14a is kept constant regardless of the load pressure, so that the volume flow to the consumer 6 is proportional to the opening cross section of the measuring orifice 14a.

When a control pressure P _{sτ is applied,} for example, to the left end face of the proportional valve 36, the valve slide 38 is shifted to the right, so that the orifice 14a for opening the connections P, Pl is opened. In the fine control range, ie in the first part of the valve spool stroke, the connection to the bypass channel connection U is still blocked. The hydraulic fluid is led via the working line 12a to the input connection Q and via the control line 18 to the left end of the control piston of the pressure compensator 16a, so that it is shifted into its control position in order to keep the pressure drop constant over the measuring orifice 14a.

The hydraulic fluid flow set in this way is then conducted via line 40, connections R, A to the work connection of consumer 6, while beitsanschluß B and the tank line 44, the hydraulic fluid from the consumer 6 is returned to the tank. Port S is closed.

When the orifice 14a is opened further, the bypass channel 32 is opened by the valve slide 38, so that the hydraulic fluid flows directly into the line 40. The volume flow to the pressure compensator 16a is reduced or even completely shut off, so that a larger volume flow is led to the consumer 6. This increase in the volume flow also leads to a drop in the system pressure when the higher-load consumer 4 has moved to a stop.

Fig. 3 shows a section through a directional valve segment through which the circuit shown in Fig. 2 is realized. The directional control valve segment has a valve plate 52 in which receiving bores for the valve slide 38, the pressure compensator 16a, two pressure limiting valves 54a, 54b and the two check or load holding valves 56a, 56b are formed. In the valve plate 52, the two working connections A, B, two control connections 58a, 58b for controlling the proportional valve 36, a pump connection P, at least one connection for the load pressure signaling line 22 and a tank connection are also provided.

The basic structure of this directional valve segment is already known from the prior art and is described, for example, in W095 / 32364 mentioned at the beginning.

The valve spool 38 has a control collar 60 in its central region, which cooperates with a web

62 of the valve bore forms the orifice 14a. In the illustration according to FIG. 3, the valve slide 38 is in its basic position by the two compression springs 41a, 41b biased in which there is no flow through the orifice 14a.

The proportional valve 36 is activated by applying a control pressure to the two control connections 58a and 58b, which are connected to the spring chamber 64a and 64b of the proportional valve 36 via control lines. In the control line between the control connections 58a, 58b and the spring spaces 64a and 64b, a nozzle with a check valve is formed, by means of which damping of the valve slide movement is possible.

The control collar 60 is provided in the region of its end faces with a plurality of control notches 64 and 66, via which pressure medium can be guided from an annular space 68 connected to the pump connection P to the input connection Q, so that the lower end face of the control piston 72 in FIG. 3 the pressure compensator 16a can be acted upon with the pressure downstream of the measuring orifice.

When the directional valve spool 38 is shifted to the right (FIG. 3), the measuring orifice 14a is formed by the interaction of the control notches 64 with the one control edge of the web 62, while the control notches 66 open the connection from the annular space 68 to the pressure compensator 16a in the event of a displacement to the left .

The input connection Q of the pressure compensator 16a is designed as an axial connection, so that the fluid pressure also acts on the lower end face 70 of the control piston 72. The output connection C is designed as a radial connection and opens into the lines 40 and 42. In these lines 40, 42, the load-holding valves 56a, 56b are arranged, which prevent a backflow from the valve spool 38 to the pressure compensator 16a and allow flow through in the opposite direction. The connection of the lines 40, 42 to the working connections A or B or the tank connection T is in each case via a directional part of the valve slide 38. That is, each working connection A, B is assigned a directional part, via which a working connection A or B with a Line 40, 42 or can be connected to the tank T.

The directional part for the connection B, which is formed on the right in FIG. 3, has three control collars 74, 76 and 78 formed in the axial distance. The control collars 76 and 78 are each provided with control notches 80 and 82, respectively, which are arranged between the control collars 76, 78 open radially recessed section.

The directional part of the valve slide 38 assigned to the working connection A is formed only by two spaced-apart control collars 84, 86. In the tax union 86 are

Control notches 88 formed, which correspond in function to the control notches 80 of the control collar 78.

At an axial distance from the right end face of the control collar 86, a plurality of oblique bores 90, which are distributed on the circumference and which are connected to a common axial bore 92, open on the outer circumference. This passes through the control collar 8 to the left end section of the valve slide 38. In the variant shown, the end stop 94 of the valve slide is screwed into the axial bore 92, so that its left end section is closed.

4 shows a detailed illustration of the valve slide 38 in the central region of this axial bore 92. Accordingly, a retaining valve is provided in the axial bore 92, the valve body 96 of which is prestressed against a valve seat 98 via a compression spring 97.

A radial bore star 100 and an oblique bore star 102 open downstream of the valve body 96. The radial bore star 100 is blocked by a web 104 of the receiving bore 103 of the valve slide 38. The oblique bore star 102 opens into the radially recessed section between the control collars 84 and 86. The valve body 96, which is biased against the valve seat 98, prevents hydraulic fluid from flowing into the axial bore 92 from the connection A. A flow in the opposite direction is practically not prevented since the compression spring 97 is weak.

The geometry of the radial bore star 100 and the oblique bore star 102 is selected such that when the valve slide 38 is shifted to the left via these stars 100, 102, the connection from the working connection A to the tank connection T can be controlled. Alternatively, control notches in the right face area of the control collar 84 could of course also be used for the control.

If a control pressure is now applied to the control connection 58a, the valve slide 38 is moved to the right in the illustration according to FIG. 3, so that the control notches 64, in cooperation with the web 62, open the connection from the pump connection P to the input connection Q of the pressure compensator .

The end face 105 of the control piston 72 located at the top in FIG. 3 is acted upon by the force of a control spring 106 and by the load pressure, which is controlled by a control edge and an angular bore 108 in the control piston 72. catch groove 110 is tapped. The pressure at the input port Q downstream of the orifice 14a deflects the control piston 72 upwards and the output port C is opened until a force equilibrium is established above the control piston 72. The load holding valve 56a is opened and the hydraulic fluid is led to the working connection A via the line 40 and the control collar 86 with the control notches 88. At the same time, the connection between the working connection B and the tank connection T is opened via the control collar 76 assigned to the working connection B and the control notches 82, so that the hydraulic fluid can flow back into the tank from the consumer. In this fine control area, the oblique bores 90 of the bypass channel 32 have not yet been opened by the control edge 107.

When the valve spool 38 is displaced further, the control edge 107 opens the bypass channel 82, so that the hydraulic fluid or at least a partial volume flow is led to the working connection A. The system pressure drops, so that the load-lower consumer 6 can be operated at a higher speed.

When the valve slide 38 is actuated in the opposite direction, the bypass channel has no effect, since the reverse flow from A to the inlet connection Q of the pressure compensator 16a is prevented by the valve body 96 resting on the valve seat 98.

In the exemplary embodiment described above, the bypass channel 32 is only assigned to the working connection A, which is required for the lifting function of the consumer. Of course, a further bypass channel can also be assigned to the other work connection B, which would then have an identical structure to the work connection described above. The diagram according to FIG. 5 shows the pressure and volume flow ratios of the above-described processes as a function of time. It is assumed that a higher load consumer, for example a spoon, is first moved to a stop. The corresponding pressure curve is shown in Fig. 5 with solid lines. Accordingly, the load pressure at this consumer increases very quickly and reaches a maximum at time t1, which is predetermined by the pump power p _sys .

After reaching this maximum pressure, a lower load consumer, for example a boom, is activated. When the proportional valve 36 assigned to this consumer is activated, the bypass channel 32 is opened in the manner described above, so that the hydraulic fluid flow Q rises to the load-lower consumer (dashed line). Due to this increase in the hydraulic fluid volume flow to the load-lower consumer, the pressure drops from the system pressure Pgys ^au ^ ^{e: Ln} lower level p *. The pressure level p * can be set by a suitable choice of the bypass duct diameter, so that the pressure drops, for example, from a pressure of 240 bar to a pressure p * of 200 bar.

At the start of the control of the lower load consumer, there is no influence on the pressure p, since the bypass channel has not yet been opened at the start of the control.

Of course, the invention is in no way limited to the fact that the bypass channel 32 is integrated in the proportional valve 36. Other solutions are also conceivable in which the bypass channel is external

Circuits is realized. An LUDV circuit for controlling at least one load-lower and one load-higher consumer is disclosed, each consumer being assigned a measuring orifice and a downstream pressure compensator for keeping the pressure drop across the measuring orifice constant. A pressure-controllable bypass channel is assigned to the pressure compensator of the lower load, via which the pressure compensator of this consumer can be bypassed.

Claims

claims

1. Hydraulic circuit for controlling at least one load-lower and one load-higher consumer (4, 6), with a variable displacement pump (2), the setting of which can be changed as a function of the highest load pressure of the consumers (4, 6), with the variable displacement pump ( 2) and each consumer (4, 6) an adjustable orifice plate (14a, 14b) with a downstream pressure compensator (16a, 16b) is provided, the control piston (72) in the closing direction of the load pressure of the associated consumer (4, 6) and in the opening direction can be acted upon by the pressure downstream of the measuring orifice (14a, 14b), characterized by

a bypass channel (32) which connects the metering orifice outlet (Pl) bypassing the associated individual pressure compensator (16a) to at least one working connection (A) for the load-lower consumer (6).

2. Hydraulic circuit according to claim 1, characterized in that the measuring orifice (14a, 14b) is formed by a proportional valve (36) through which the working connection (A, B) with the pump connection (P) or a tank (T) can be connected and that the bypass channel (32) can be opened as a function of the valve slide position of the proportional valve (36).

Hydraulic circuit according to claim 2, characterized in that the bypass channel (32) is formed in the valve slide (38) and can be opened by a control edge of the proportional valve (36). 00/00747. -ig.

4. Hydraulic circuit according to one of the preceding claims, characterized in that a non-return valve (96, 97, 98) is arranged in the bypass channel (32), which prevents hydraulic fluid flow from the consumer (6) to the metering orifice (14a).

5. Hydraulic circuit according to one of claims 2 to 4, characterized in that the proportional valve (36) has two working connections (A, B) for the consumer (6), and that each working connection (A, B) has a bypass channel (32) assigned.

6. Hydraulic circuit according to one of claims 2 to 5, characterized in that the bypass channel (32) is opened only after a predetermined stroke of the valve slide (36).

7. Hydraulic circuit according to one of claims 2 to 6, characterized in that the valve slide (38) has an approximately centrally arranged, the measuring orifice (14a) forming speed part and two directional parts, via which the hydraulic fluid from the output port (Q) Pressure compensator (16a) can be guided to a work connection (A, B) or from the other work connection (A, B) to a tank connection (T), the bypass channel (32) extending from the speed part to one of the direction parts.

8. Hydraulic circuit according to one of claims 4 to 7, characterized in that the bypass channel

(32) on the one hand via oblique bores (90) in the area of the speed section and on the other hand via a radial bore star (100) and / or an oblique bore star (102) downstream of the check valve (96, 97, 98) in the area of a directional section.

9. Hydraulic circuit according to one of the preceding claims, characterized in that the variable pump (2) is pressure and power controlled.