EP0563514B1 - Load-independent hydraulic driving apparatus for implements of construction machines - Google Patents

Abstract

A load-independent, hydraulic driving apparatus for implements of construction machines, agricultural machines, industrial trucks and the like is provided with a central, hydraulically adjustable, load-sensing variable-displacement pump (1) for delivering pressure medium from a pressure-medium reservoir (2) to the working cylinders of the implements, control spools (6) allocated to the working cylinders for activating the working cylinders, and a hydraulic load-sensing pressure-indicating device (5). A control pressure which is clearly above the applied load-sensing load pressure can be admitted to an actuating member (4) of the variable-displacement pump (1) independently of the load-sensing indicating device (5). The driving apparatus can thus adjust the working cylinders in a quick traverse with increased rapidity. <IMAGE>

Description

The invention relates to a hydraulic drive device for work equipment on construction machines, agricultural machinery, industrial trucks and the like which is independent of the load pressure. According to the preamble of patent claim 1.

Such drive devices are used as mobile hydraulics, for example, in degrees and are used there to actuate the share, the front plate, the rear ripper and other functions.

The "load pressure independent" attribute relates to the adjustment speed of the individual work tools, which acts on the work tools independently of the load pressure of the substrate being treated. This independence is achieved by a so-called load sensing system, which essentially comprises a central, hydraulically adjustable adjusting pump with a hydraulically actuated adjusting element for regulating the delivery pressure and a hydraulic pressure signaling device for transmitting the load pressure applied to a controlled working cylinder to the adjusting element of the adjusting pump. This pressure signaling device transmits the pending load pressure to the actuator of the variable displacement pump, which acts on the corresponding control slide with a delivery pressure which is a design-dependent, constant differential pressure above the pending load pressure. If several working cylinders are actuated at the same time, the highest load pressure is reported to the actuator of the variable pump via corresponding shuttle valves in the branches of the pressure signaling device.

Due to the control of the corresponding control spool with a constant differential pressure above the load pressure applied to the working cylinder, the working device actuated by the working cylinder is moved at a substantially constant speed which is dependent on the position of the control spool, which considerably facilitates the operation of such working devices. In addition, due to the load sensing system, the variable displacement pump only has to be operated at partial load with a comparatively low flow rate requirement, as a result of which no flow rate-dependent losses occur and the efficiency of the hydraulic drive device is increased.

Load pressure-independent, hydraulic drive devices with load sensing systems have the disadvantage, due to the associated adjustment speed of the work tools, that faster movements of the work tools, such as those when positioning or returning, during empty strokes or when evading, especially of the work tools, are desirable or necessary in dangerous situations , cannot be carried out. The system-related, design-based adjustment speed of the work tools is not sufficient for a quick reaction to non-operational occurrences, which is e.g. also has an adverse effect on the working speed when avoiding or bypassing obstacles such as shafts, curbs etc.

EP-A 439 621 discloses a device for loading the cylinders of work machines, with which the purpose is achieved of being able to reduce the oil flow to the cylinders for fine work. For fine work, the control described for the cylinders should respond so that an extremely small amount of oil is supplied to them. The connection of the pump line to the pump controller is in this case analogous to the subject matter of the invention, but the objective of increasing the speed of the working tools is not desired and is also not achievable.

Proceeding from this, the object of the invention is to develop a hydraulic drive device of the type mentioned at the outset that is independent of the load pressure in such a way that a significantly higher adjustment speed of the implements is achieved independently of the load sensing system. So a so-called "rapid traverse" should be created for the tools.

This object is achieved by the features specified in the characterizing part of the first claim.

Accordingly, the invention is embodied by the basic idea that the control element of the variable displacement pump can be acted upon with a control pressure which is significantly above the load pressure present, independently of the pressure signaling device. The variable displacement pump thus works in this operating state with a significantly higher delivery pressure than the pressure given by the load sensing system, so that a controlled working cylinder is moved with a correspondingly significantly increased adjustment speed. The higher control pressure can be taken from any pressure accumulator or generator available in the hydraulic system. The drive device according to the invention consequently has a "rapid traverse", with the help of which the machine operator moves to the entry can react quickly to non-operational incidents mentioned and position or return the tools in a shorter time and in particular can perform rapid idle strokes.

If the actuator of the variable displacement pump is operated in rapid traverse with a control pressure of the level specified in claim 2, the variable displacement pump delivers abruptly within its actuating time with maximum nominal pressure, so that a maximum adjustment speed of the controlled implements is achieved.

As soon as the control valve in the feedback line between the delivery output of the variable displacement pump and its actuator has been opened when the rapid traverse is switched on, the delivery pressure of the pump, which is higher by a design-dependent differential pressure, is placed on the actuator of the pump, which in turn leads to the delivery pressure of the Pump rises. The feedback means that the pump is set to the maximum nominal pressure within its actuating time, which is in the range of fractions of a second.

If the control valve in the feedback line is closed, the load pressure supplied by the pressure signaling device is again present at the control element of the feed pump, so that the hydraulic system again works with load pressure-independent adjustment speed of the working cylinders predetermined by the position of the control slide.

The blocking element in the pressure signaling line prevents the pressure medium under high pressure from flowing to the pressure medium tank against the pressure signaling direction of the load sensing system when the rapid traverse is actuated.

The blocking element can advantageously be a check valve or a throttle. Although a throttle has no preferred direction in its blocking effect, it is suitable as a blocking element acting counter to the pressure signaling direction. When the load sensing system is effective - that is, when the rapid traverse is inactive - the throttle practically does not resist the pressure transmission within the pressure signaling line without any appreciable flow of pressure medium. When rapid traverse is activated, the throttling effect is sufficient to prevent any significant outflow of pressure medium in the opposite direction in the pressure signaling direction.

The further claim characterizes an advantageous further development of drive devices in which pressure scales are provided in connection with the load sensing system, each of which is connected upstream of one or more control slides of the working cylinders combined into blocks. These pressure compensators, the control input of which is each acted upon by a corresponding load pressure, have the function that the pressure medium is only passed on at its outlet with a pressure which is higher than the load pressure by a design-related differential pressure. If two working cylinders are actuated at the same time, the variable pump delivers pressure medium with a delivery pressure that is higher by the design-related differential pressure than the higher of the load pressures applied to the two working cylinders, but the action of the pressure compensator that acts on the lower load pressure Working cylinder is assigned, the delivery pressure of the variable pump is reduced to a value which is higher than the load pressure by the design-related differential pressure. The delivery flow is therefore supplied to both control slides with different pressures, which are both, however, at the same differential pressure above the respective load pressure. This means that both working cylinders are actuated at the adjustment speed preselected via the position of the corresponding control slide.

From the above explanation it is clear that without further measures, the rapid traverse would be ineffective in hydraulic systems with pressure compensators, since due to their intended function they would reduce the increased delivery pressure of the variable pump to an admission pressure, which in turn is only above the load pressure by the design-related differential pressure. So there would be no change in the adjustment speed as it is intended in rapid traverse.

In order to avoid this, the control inputs of the pressure compensators can also be acted upon by the increased control pressure acting on the control element of the variable displacement pump. This practically simulates a load pressure that is higher than the actual load pressure at the control inputs, as a result of which the pressure compensators remain ineffective and allow the full delivery pressure of the variable displacement pump to pass through the corresponding control slide to the working cylinders. The latter are thus actuated with the control slide corresponding to the maximum flow rate of the variable displacement pump when the control slide is fully open.

Further features, details and advantages of the invention can be found in the following description, in which exemplary embodiments are explained in more detail with reference to the accompanying drawings.

FIGS. 1 to 3 show circuit diagrams of three drive devices according to the invention in different embodiments.

FIG. 1 shows a hydraulic drive device with a load sensing system that is independent of the load pressure and is used, for example, to actuate the working cylinders (not shown) on the blade of a grader. The drive device comprises a central variable displacement pump 1, which pumps the pressure medium from a tank 2 as a pressure medium reservoir into the flow line 3 with a controllable delivery pressure p _F. The variable displacement pump 1 is of a conventional type. To regulate the delivery pressure, it has a hydraulically actuated actuator in the form of a pressure delivery flow regulator 4, the control input of which is connected to the hydraulic pressure signaling device 5. The flow line 3 branches to the three identically constructed, conventional control slides 6, which are actuated by mechanical devices. Hydraulic actuation is usually also possible.

On the supply side, a return line 7 opening into the tank 2 extends from this control slide 6. Two connecting lines 8, 9 continue to lead from the control slide 6 to the corresponding (not shown) working cylinders. These connecting lines 8, 9 can be connected to the supply or return line 3 or 7 depending on the desired actuation direction of the working cylinder. Each control slide 6 is also provided with a pressure signaling connection 10, via which the load pressure present on the consumer side can be tapped when the control slide 6 is actuated. The branched pressure signaling line 11 of the pressure signaling device 5 is connected to the pressure signaling connections 10 of the control slide 6, which line leads to the pressure flow control device 4. At the branching points of the pressure signaling line 11, shuttle valves 12 are installed, with the aid of which the highest load pressure p _{LS present at} the pressure signaling connections 10 of the individual control spools 6 is fed to the pressure-flow regulator 4 as control pressure while simultaneously actuating several control spools 6. The pressure signaling line 11 is also connected to the return line 7 via a tank connecting line 13.

The load sensing systems described above with pressure signaling devices 6 are common St.dT and are widely used in mobile hydraulic devices of construction machines, agricultural machinery, Industrial trucks and the like. Used. The unit consisting of variable pump 1, pressure-flow controller 4 and pressure signaling device 5 has the function that the delivery pressure p _{F of} the variable pump 1 is set so that it is a constant, design differential pressure Δ p _LS above the highest of the load sensing pressure signaling device 5 Pressure flow controller 4 reported load pressure p _LS is. Δ p _LS is in the order of 10 to 25 bar. The adjustment speed proportional to the root of the pressure difference Δ p _LS is therefore also constant and cannot be increased per se due to the system.

In order to have a rapid traverse available, a feedback line 14 is provided between the flow line 3 and the pressure detection line 11. In this feedback line 14 there is a control valve 15 for opening and closing this line, which is actuated electromagnetically. Manual, hydraulic or pneumatic actuation is also possible. With the help of the feedback line 14 and the control valve 15, the delivery output of the variable displacement pump 1 can be connected directly to the pressure delivery flow controller 4 of the pump 1. As soon as the control valve 15 is switched to passage, the delivery pressure p _{F of} the variable pump 1 rises within a fraction of a second within its regulating time to the maximum pressure that can be reached according to the pressure regulator setting, which is generally the maximum nominal pressure of the pump 1. This automatic setting is based on the effect of the pressure flow regulator 4 of the variable displacement pump 1, namely that the delivery pressure p _{F is} always built up by a design differential pressure Δ p _LS higher than the load pressure p _LS reported on the pressure flow regulator 4.

A check valve 16 is so that when the rapid traverse is actuated and the control valve 15 is switched through, no pressure medium can flow via the pressure reporting line 11, which is pressurized with the maximum system pressure, depending on the random position of the shuttle valves 12, via one of the control slide 6 or the tank connecting line 13 to the tank 2 installed in the pressure signaling direction D in front of the junction point of the feedback line 14 in the pressure signaling line 11, the blocking direction of the check valve 16 pointing counter to the pressure signaling direction D. If the rapid traverse via the control valve 15 is not actuated, the feedback line 14 and in particular the check valve 16 are ineffective, since in the load sensing pressure message in the pressure reporting line 11 an oil flow only flows from the control slide valve 6 to the pressure / flow rate controller 4.

In the exemplary embodiment of a drive device according to the invention according to FIG. 2, two separate control slide blocks 17, 18 are each connected in parallel hydraulically Control slide 6.6 'provided. The blocks 17, 18 are each constructed essentially in accordance with the exemplary embodiment according to FIG. 1. The same components are provided with the same reference symbols (if necessary with the addition of an apostrophe). To avoid repetitions, only the differences from the exemplary embodiment according to FIG. 1 are explained:

In the first place, each control slide block 17, 18 is provided on the input side with a pressure compensator 19, 19 'in the branching supply line 3. These pressure compensators 19, 19 'have control inputs which are acted upon by the load pressure present on the outlet side. The task of the pressure compensators 19, 19 'is to throttle the flow rate through the feed line 3 to such an extent that the pressure on the output side of the pressure compensator 19, 19' is a constant differential pressure above the load pressure present at its control input, regardless of the input pressure. This differential pressure is preferably selected in accordance with the differential pressure of the load sensing system.

The pressure signal line 11 is also branched towards the two control slide blocks 17, 18 (line branches 11, 11 ') in accordance with the flow line 3, a further shuttle valve 20 being arranged at the branching point.

Compared to the exemplary embodiment according to FIG. 1, the components responsible for the rapid traverse function are varied. The control valve is a branch valve 21 which divides the feedback line 14 into two partial branches 22, 23 on the part of the pressure-flow regulator 4. These each open into the two branches 11, 11 'of the pressure reporting line in relation to the pressure reporting direction D before the shuttle valve 20. In addition, stub lines 24, 24 'are provided from the two branches 11, 11' of the pressure signal line to the control inputs of the pressure compensator 19, 19 '. In relation to the pressure signal direction D before the branch points of the branch lines 24, 24 ', check valves 16, 16' are again provided.

When rapid traverse is not activated, the feedback line 14, its branches 22, 23 and the check valves 16, 16 'remain ineffective. Because of the shuttle valves 12, 12 ', the highest load pressure p _{LS is reported} to the control spool 6 or 6' in the two branches 11, 11 'of the pressure signaling line. Because of the shuttle valve 20, only the higher of these two pressures is fed to the pressure-flow regulator 4. As a result, the delivery pressure p _{F of} the variable displacement pump 1 is set to a value which is higher by the design-related differential pressure Δ p _LS than the highest load pressure p _LS . Is this delivery pressure correct? For example, after a higher load pressure applied to a spool 6 of the spool block 17, the pressure compensator 19 in this spool block 17 remains ineffective. The pressure compensator 19 'in the other control slide block 18, however, only a small load pressure is reported via the spur line 24', so that it takes action and reduces the excessive delivery pressure present on the pressure compensator 19 'on the input side so that the output side to the corresponding control slide 6' only there is a delivery pressure that is higher than the load pressure by the design-related differential pressure. With simultaneous activation of two control slides 6, 6 'in the two control slide blocks 17, 18, the adjustment speeds of the working cylinders supplied with them, which are preselected via the position of the control slides, are ensured.

When the rapid traverse is actuated by a corresponding actuation of the branch valve 21, the delivery pressure of the variable displacement pump 1 is passed to the pressure delivery line controller 4 via the feedback line 14, the two branch branches 22, 23 and the section of the pressure detection line 11 leading to the pressure delivery flow controller 4, so that the variable displacement pump 1 - as described - brings the maximum delivery pressure within fractions of a second within its control time. On the other hand, this pressure is supplied via the two branches 11, 11 'of the pressure signal line and the stub lines 24, 24' to the two pressure compensators 19, 19 ', so that these become inoperative, since from the supply line 3 to the pressure compensators 19, 19' as well only the delivery pressure can act. The flow rate thus passes the pressure compensators 19, 19 'without throttling and thus without any significant pressure loss.

The embodiment shown in FIG. 3 essentially corresponds to that shown in FIG. 1 with the difference that a pressure compensator 19 is connected upstream of each control slide 6. Otherwise, the same reference numerals designate the same components, so that a repeated description of these components is unnecessary.

For the rapid traverse function, according to the exemplary embodiment in FIG. 2, a branch valve 21 is provided in the feedback line 14, which divides the feedback line 14 into two partial branches 22, 23 on the part of the pressure-flow regulator 4. Partial branch 22 opens directly into the pressure signaling line 11 of the signaling device 5. The second partial branch 23 branches via branch lines 25 to the control inputs of the pressure compensators 19. Furthermore, a connecting branch 26 to the spool-side branches of the pressure signaling line 11 is provided between the branch lines 25. Here, based on the pressure signaling direction D, both in the branch lines 25 and in the connecting branches 26 are each before the junction of the connecting branches 26 installed in the branch lines 25 check valves 27, 28, the blocking direction of which points counter to the pressure signaling direction D.

When the rapid traverse is actuated by a corresponding actuation of the branch valve 21, the output of the variable displacement pump 1 is in turn fed via the feedback line 14, its partial load 22 and the pressure signaling line 11 to the pressure flow regulator 4 of the variable displacement pump 1, so that the latter delivers the maximum delivery pressure. This is placed simultaneously on the branch 23 and the branch lines 25 on the control inputs of the pressure compensators 19, so that these - as already described above - become ineffective. The check valves 28 are provided in the connecting branches 26 so that the maximum delivery pressure cannot escape via the connecting branches 26 to the tank 2.

If the rapid traverse is not actuated, the check valves 27 shut off the branch lines 25 against the pressure detection direction D, so that the load pressure applied to the control slide 6 is reliably fed to the control inputs of the pressure compensators 19.

Finally, a brief design example should be given of which higher adjustment speeds can be achieved in rapid traverse:

mit v:

Eilgang-Verstellgeschwindigkeit

mit v_LS:

Loadsensing-Verstellgeschwindigkeit.

The larger pressure difference Δ p _H (difference between maximum pressure and load pressure) compared to the differential pressure Δ p _LS causes the higher adjustment speed according to the relationship: $${\text{v: v}}_{\text{LS}}\text{=}\sqrt{{\text{Δ p}}_{\text{H}}{\text{: Δ p}}_{\text{LS}}}$$

with V:

Rapid traverse adjustment speed

with v _LS :

Load sensing adjustment speed.

The flow rate passing through the slide and thus the adjustment speed of the working cylinder decrease with increasing consumer pressure in accordance with the decreasing Δ p _H. At a pressure difference of Δ p _H = Δ p _LS , the flow rate ultimately becomes as large as the flow rate in load sensing mode. According to the task, however, the rapid traverse gearshift should serve to quickly position the implements or quickly out of danger zones to move, usually only lower load pressures are present, so that the theoretically achievable adjustment speed is also approximately achieved in these cases.

• Loadsensing-Differenzdruck: Δ p_LS = 14 bar
• Maximaler Pumpenförderdruck: p_max = 180 bar
• Nennförderstrom bei Loadsensing-Betrieb: Q_LS = 40 l/min
• Maximaler Förderstrom der Pumpe: Q_max = 103,5 l/min

If the maximum flow is passed through the spool valve before the maximum pump pressure is reached, the maximum speed remains unchanged up to a certain load pressure, as the following design example shows:

• Load sensing differential pressure: Δ p _LS = 14 bar
• Maximum pump delivery pressure: p _max = 180 bar
• Nominal flow rate in load sensing mode: Q _LS = 40 l / min
• Maximum flow of the pump: Q _max = 103.5 l / min

Required pressure difference at the spool valve to allow Q _{max to} flow through: $${\text{p}}_{\text{d}}{\text{= (103.5: 40)}}^{\text{2nd}}{\text{x Δ p}}_{\text{LS}}\text{= 94 bar.}$$

The maximum speed can thus be maintained up to a load pressure pL of: $${\text{p}}_{\text{L}}{\text{= p}}_{\text{Max}}{\text{- Δ p}}_{\text{d}}\text{= 86 bar.}$$

This means that the maximum adjustment speed only begins to decrease at load pressures of more than 86 bar. In the design example, this is 2.6 times the nominal speed in load sensing operation, because $${\text{v}}_{\text{Max}}{\text{= (Q}}_{\text{Max}}{\text{: Q}}_{\text{LS}}{\text{) xv}}_{\text{LS}}{\text{= (103.5: 40) xv}}_{\text{LS}}{\text{= 2.6\%}}_{\text{LS}}\text{.}$$

Claims (7)

1. A load pressure-independent hydraulic drive device for tools on construction machinery, agricultural machines, industrial vehicles and similar machines, in particular for the blade, front bowl, rear-mounted ripper or similar devices on motor graders, having

   - a central, hydraulically adjustable, variable displacement pump (1) for delivering pressure medium from a pressure medium reservoir (tank 2) to the working cylinders of the tools, which pump (1) comprises a hydraulically operable adjusting element (pressure delivery flow regulator 4) for regulating the delivery pressure (p_F),

   - spool valves (6, 6') associated with the working cylinders for controlling the working cylinders, and

   - a hydraulic pressure-signalling device (5) integrated in a pressure-signalling line (11) for transmitting the load pressure (p_LS) present at a controlled working cylinder to the adjusting element (pressure delivery flow regulator 4) of the variable displacement pump (1), which pump pressurizes the corresponding spool valve (6, 6') with a delivery pressure (p_F) which lies above the existing load pressure (p_LS) by a constant differential pressure (Δ p_LS) imposed by the design,

   characterized in that the delivery output of the variable displacement pump (1) is arranged to be connected to the adjusting element (pressure delivery flow regulator 4) thereof by way of a feedback line (14), which pressurizes the adjusting element (4) with a control pressure clearly above the existing load pressure (p_LS), in that in the feedback line (14) there is arranged a control valve (15), a branch valve (21) for opening and closing the feedback line (14), and in that the pressure-signalling line (11) leading to the adjusting element (4) of the variable displacement pump (1) is provided with a blocking member (non-return valve 16, 16') acting against the pressure-signalling direction (D).
2. A drive device according to claim 1, characterized in that the control pressure for the adjusting element (pressure delivery flow regulator 4) of the variable displacement pump (1) corresponds at least to the maximum nominal pressure (p_max) of the pump (1) minus the constant differential pressure (Δ p_LS).
3. A drive device according to claims 1 and 2, characterized in that the blocking member is a non-return valve (16, 16').
4. A drive device according to claims 1 and 2, characterized in that the blocking member is a throttle.
5. A drive device according to one of claims 1 to 4, having at least two pressure balances (19, 19'), each of which is connected in series with one or more spool valves (6, 6') of the working cylinders, characterized in that the control inputs of the pressure balances (19, 19') are pressurizable with the control pressure acting on the adjusting element (pressure delivery flow regulator 4) of the variable displacement pump (1).
6. A drive device according to one or more of claims 1 to 5, characterized in that the control valve in the feedback line (14) is in the form of a branch valve (21) having an input and two outputs, at least one output being arranged to be connected to the control inputs of the respective pressure balance (19, 19') and at least one output being arranged to be connected to the adjusting element (pressure delivery flow regulator 4) of the variable displacement pump (1).
7. A drive device according to claim 6, having several spool valves (6), each of which has a pressure balance (19) connected in series, each spool valve (6) being connected on the load side by way of a pressure-signalling line (connecting branches 26) to the control input of the pressure balance (19) associated therewith, characterized in that each control input of the pressure balance (19) is connected by way of a connecting line (branch line 25, part-load 23) to the one output of the branch valve (21), and each pressure-signalling line (connecting branches 26) and each connecting line (branch lines 25) is provided with a respective blocking member (non-return valves 27, 28) acting in a direction against the pressure-signalling direction (D).

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