DE102021210054A1 - Energy efficient electric-hydraulic control arrangement - Google Patents

Abstract

Hydraulic drive (1) for a mobile work machine, preferably an excavator, with a hydraulic machine (4) that can pivot, which is hydraulically connected to first consumer cylinders (44), drives them and can regenerate potential energy when the first consumer cylinders (44) retract. Furthermore, the hydraulic drive (1) has a hydraulic pump (2), which is hydraulically connected to further consumer cylinders (90) and drives them. The pivotable hydraulic machine (4) is connected to the first consumer cylinders (44) via a first control circuit (5) and the hydraulic pump (2) is connected to the other consumer cylinders (90) via a second hydraulic control circuit (84), and the control circuits (5, 84 ) are hydraulically separated from each other.

Description

technical field

The present invention relates to a hydraulic drive for a mobile work machine, preferably an excavator with a hydraulic pump and a hydraulic machine that has the option of recuperation.

State of the art

From the DE 10 2014 216 031 A1 a hydrostatic drive for the recuperation of hydraulic energy is known. The EP 3 358 201 A1 discloses a hydraulic energy regeneration apparatus having a hydraulic pump and a hydraulic generator.

From the EP 2 738 397 B1 is a construction machine with a variable displacement pump and a combined hydraulic pump and generator, ie a combined hydraulic machine is known. This combined hydro machine can recover energy but also pump a working fluid. A total pressure level for all hydraulic consumers is established in the named drives. This entire pressure level is specified by the consumer with the highest load pressure. This results in power losses that result from high pressure differences that have to be throttled at individual pressure compensators, for example. These throttling losses occur mainly in part-load operation and when consumers are (simultaneously) operated in parallel and are due to the design as a single-circuit system.

Summary of the Invention

The invention is therefore based on the object of providing an improved hydraulic drive, in particular one that is as energy-efficient as possible, for lifting and lowering a load. Furthermore, as few components as possible should preferably be required for the hydraulic drive.

This problem is solved by a hydraulic drive with the features of claim 1. Advantageous developments of the invention are the subject matter of the appended dependent claims.

The invention accordingly relates to a hydraulic drive for a mobile electrified working machine, preferably an excavator. The hydraulic drive has a pivotable hydraulic machine that is provided and designed to drive first consumer cylinders and to regenerate potential energy when the first consumer cylinders are retracted/retracted. The hydraulic drive also has a hydraulic pump that drives further/other consumer cylinders. According to the disclosure, the hydraulic machine is connected to the first consumer cylinders via a first control circuit and the hydraulic pump is connected to the further consumer cylinders via a second control circuit, and the two control circuits are hydraulically separated from one another.

The hydraulic drive has the two hydraulic machines/hydraulic pumps. The first hydraulic machine/hydraulic pump is hydraulically connected to a first consumer, preferably a boom of the excavator, and raises and lowers the first consumer. As the boom lowers, potential energy is converted to kinetic energy. A flow of fluid flows from a consumer cylinder back to the first hydraulic pump. This kinetic energy can be regenerated by the first hydraulic pump, since the first hydraulic pump is capable of pivoting or mooring. That is, the first hydraulic pump can change its swivel angle to such an extent that it functions as a hydraulic generator.

In other words, the hydraulic pump capable of pivoting is a combined hydraulic pump/generator (hydraulic machine). A drive shaft of the hydraulic machine does not have to change its direction of rotation for this. The first hydraulic pump is therefore a combination of a hydraulic pump and a hydraulic motor and can run in one conveying direction, for example to pump a fluid, but can also be driven by a fluid in another conveying direction as a hydraulic generator. The second hydraulic pump is intended and designed exclusively for operating additional consumers. The other consumers are, for example, a stick, a bucket and a retarder of the excavator. Further consumers can be, for example, drives, adjustable booms, shields and/or optional consumers. The first hydraulic pump is connected to the first consumer via a first hydraulic control circuit. Likewise, the second hydraulic pump is connected to the other consumers via a second hydraulic control circuit. The two control circuits are separated from one another in such a way that different pressure levels are possible in the respective control circuits.

The separate control circuits enable energy-efficient distribution through multiple working fluid flows, provided by multiple hydraulic pumps, to the main and auxiliary linear consumers. Due to the hydraulic machine that can be pivoted, the potential energy when lowering the consumer cylinder can preferably be used without reversing the direction of rotation Drive shaft that drives the pump can be recovered. Energy recovery when lowering the boom or retracting the consumer cylinders under active loads is made possible. The separate control circuits make it possible to apply different pressure levels in the respective control circuits.

• • energieeffiziente Steueranordnung zum Betrieb einer mobilen Arbeitsmaschine/ eines elektrifizierten (Mobil)Baggers
• • der elektrifizierte (Mobil)Bagger kann lokal emissionsfrei arbeiten
• • Adressierung der relevanten bzw. wichtigsten Einspar- und Rückgewinnungspotentiale hinsichtlich des energetischen Verhaltens und somit kosteneffizienter Aufbau möglich
• • Nutzung von Rohrbruchsicherungsventilen als Leckagekompensation beim regenerativen Absenken des Verbraucherzylinders
• • Anordnung eines Niederdruckspeichers zum Schutz der durchschwenkfähigen Pumpe durch Verhinderung von zu niedrigen Drücken unter z. B. einem bar im Druckbereich

The present disclosure has the following advantages.

• • Energy-efficient control arrangement for operating a mobile work machine/an electrified (mobile) excavator
• • the electrified (mobile) excavator can work locally without emissions
• • Addressing of the relevant or most important savings and recovery potentials with regard to the energetic behavior and thus cost-efficient construction possible
• • Use of pipe rupture safety valves as leakage compensation during regenerative lowering of the consumer cylinder
• • Arrangement of a low-pressure accumulator to protect the pivotable pump by preventing pressures that are too low, e.g. B. a bar in the pressure range

According to a further feature of the invention, a fluid flow of the hydraulic pump is added to a fluid flow of the hydraulic machine by a summing valve and vice versa. The two control circuits can be switched together by the summation valve. The volume flows of the two pumps are added as required to form a total volume flow in one of the two consumer circuits. The fluid flow of the hydraulic pump is switched on when the maximum volume flow that can be pumped by the respective hydraulic machine is not sufficient to achieve the required speed of the consumer or consumers. Due to the possibility of adding the fluid flows, the individual hydraulic pumps and components can be made smaller. This leads to reduced losses due to leaks and pressure losses due to better volumetric flow utilization, especially with partial load operation.

In a working position, the summation valve separates the two control circuits. In a working position of the summation valve, a volume flow of the hydraulic pump can be applied to a pump line of the hydraulic machine and vice versa.

According to a further feature of the invention, the pressure level in the individual control circuits is different from one another. Depending on the load situation of the consumers, different pressures can be present in the first control circuit and in the second control circuit, which is only made possible by the separation of the control circuits. The division into different pressure levels makes it possible to reduce throttling losses, especially when several hydraulic pumps and consumers are operated in parallel, which are caused by the fact that different consumers require different load pressures and an individual control circuit has to be based on the consumer with the highest load pressure. For the other consumers, the load pressure has to be throttled away, which leads to losses.

According to a further feature of the invention, the pivotable hydraulic machine regenerates the recovered energy mechanically via the common shaft directly to the hydraulic pump. The recovered energy is transferred directly to the hydraulic pump without intermediate electrical storage. As a result, the recovered energy can be transmitted without losses in the conversion of mechanical energy into electrical energy and without losses in electrical intermediate storage.

According to a further feature of the invention, the pivotable hydraulic machine converts the recovered energy into electrical energy. Kinetic energy when retracting the consumer cylinder is converted into electrical energy by the pivoting hydraulic machine. The electrical energy is stored in an accumulator that supplies the electric motor(s) of the hydraulic pump. This is particularly beneficial in the case when the recovered energy is not required directly by the hydraulic pump, but only later.

According to a further feature of the invention, the pivotable hydraulic machine is connected to a bottom side of the consumer cylinder via a pump line with an (integrated) main control slide and (integrated) individual pressure compensator. The consumer cylinders are connected to the pivotable hydraulic machine via a regeneration line/pump channel. The fluid flow of the second hydraulic pump is introduced into the pump line via the summation valve when required.

The flow of fluid from the hydraulic machine to the bottom side of the consumer cylinder is controlled by the main control slide valve and the individual pressure compensator. When the consumer cylinders are lowered, a flow of fluid occurs from the consumer cylinders to the hydraulic machine that can pivot. This fluid flow is dependent on the functioning of the swiveling Hyd raulikmaschine converted into electrical energy as a hydraulic generator. If the volume flow delivered by the hydraulic machine is not sufficient to set the required (maximum) speeds of the consumer cylinders, the fluid flow of the additional hydraulic pump can be switched on through the summation valve.

When the consumer is lowered regeneratively, a flow of fluid from the bottom side of the consumer cylinder via the pump channel to the hydraulic machine capable of pivoting is made possible. The speed of the regenerative lowering can be controlled by the swiveling hydraulic machine.

According to a further feature of the invention, the hydraulic drive has a low-pressure accumulator that provides a minimum pressure that prevents the hydraulic machine that can swivel through from running in an area with too low a pressure.

Rapid changes in load and direction of movement can lead to a pressure drop on the hydraulic machine, which is harmful to the hydraulic machine. This can be avoided by an additional volume flow from the low-pressure accumulator. The low-pressure accumulator is filled by increasing the preload of a return channel in a main valve block of the hydraulic drive. For this purpose, both the return quantities of the consumers and an additional feed by the hydraulic pump of the second control circuit via a pressure relief valve (e.g. unloading valve when the system is designed as an LUDV system) can be used.

Pressures that are too low can damage the pivotable hydraulic machine. In order to prevent this, the hydraulic drive has the low-pressure accumulator, which conveys working fluid into the pump channel via a check valve. This is necessary in particular when the user requests rapid changes in load and/or direction of movement. An example of this is when the user commands a lowering in the opposite direction immediately after lifting the consumer at high speed and the hydraulic pump swings out in the negative direction. Due to the consumer's inertia, it remains in its speed direction, but the hydraulic machine begins to suck a volume flow out of the pump channel. The same behavior occurs when the shovel hits the ground during the lowering process. This can lead to a massive drop in pressure at the pump outlet, which is not designed for pressures below e.g. 1 bar. Damage can result from this, which can be avoided by the additional flow of working fluid from the low-pressure accumulator.

According to a further feature of the invention, the hydraulic drive has pipe rupture safety valves which, controlled electrically, set an opening cross section by which a volume flow from the bottom side of the consumer cylinder into the pump channel is controlled.

In regenerative lowering, the burst check valves can set an opening area that controls the amount of working fluid flowing from the bottom side of the consumer cylinders to the hydraulic machine. This allows the regenerative lowering of the boom to be controlled more precisely than would be possible with the hydraulic machine. The pipe rupture safety valves thereby allow fine control during regenerative lowering of the boom.

The pipe rupture safety valves are controlled independently of the main control slide. All proportional 2/2-way valves with a non-return function in one direction and a flow function in the other direction can be used as pipe rupture safety valves. The pipe rupture safety valves allow a fine control range, which is necessary for crane work, to be mapped up to a certain speed. From the specified speed, the pipe rupture safety valve opens completely and the speed during regenerative lowering is only measured by the swiveling hydraulic machine, which works as a hydraulic motor and can return power to the drive shaft.

In addition, the pipe rupture safety valve in the fine control area can be used to compensate for all internal leaks in the lowering path of the hydraulic drive in order to enable a steady increase in speed from the rest position and thus avoid an offset in the speed profile.

The invention ensures the lowering movement in such a way that component and system-related leaks are compensated. The component and system-related leaks occur, for example, in a flow regulator in a load pressure reporting channel or a relief line for pressure relief, a first bypass valve in the pump channel to the pivoting hydraulic machine and through leaks in the pivoting hydraulic machine.

According to a further feature of the invention, the hydraulic drive has the first bypass valve, which has a hydraulic connection between blocks and opens the pump channel and the bottom sides of the consumer cylinders.

In the classic lowering of the first consumer, the first bypass valve blocks the connection between the pump channel and the bottom side of the consumer cylinder in order to prevent an increase in pump pressure and an uncontrolled lowering of the consumer. With classic lowering, pressure from the hydraulic machine is applied to the piston sides of the consumer cylinders. Working fluid from the bottom sides of the consumer cylinders is relieved into the return passage. If the connection from the bottom sides to the pump channel were open, the bottom sides would also be relieved through the pump channel. This would increase the pump pressure and as a result be relieved via the relief line and a throttle. This would result in an uncontrolled lowering of the consumer. This uncontrolled lowering is prevented by the fact that the first bypass valve blocks the hydraulic connection between the bottom sides of the consumer cylinders and the pump channel during classic lowering.

According to a further feature of the invention, the hydraulic drive has a second bypass valve which establishes a connection between the bottom sides of the consumer cylinders and the piston sides of the consumer cylinders.

The second bypass valve increases the energy efficiency of the overall arrangement and ensures that the piston side of the consumer cylinder is filled. The second bypass valve also achieves an increase in the load pressure and a reduction in the volume flow, which must be conveyed from the piston side to the hydraulic machine capable of mooring/slewing through during the regenerative lowering of the boom. Due to the resulting reduction in pump volume flow, the hydraulic machine, including the electric motor driving the hydraulic machine, can be operated at a lower speed than would be possible with the entire boom volume flow, which leads to lower mechanical losses. The associated adjustment of the swivel angle of the hydraulic machine to almost the maximum value also helps to further minimize the friction and thus the mechanical losses.

The second bypass valve results in two advantages. The piston side of the consumer cylinder can be filled without a volume flow from the hydraulic pump. Furthermore, the volume flow that has to flow back from the bottom side of the consumer cylinder to the hydraulic machine decreases. This allows the speed to be reduced and the displacement of the hydraulic machine to be set to a maximum value. In addition, the pressure losses in the pump channel caused by the volume flow are reduced.

According to a further feature of the invention, the hydraulic pump enables a flow of fluid to the piston side of the consumer cylinder via the summation valve during the recuperative lowering of the first consumer. As a result, the recuperative lowering can be accelerated by an activated active lowering and the filling of the piston sides of the cylinders can be ensured if the regenerative lowering is carried out without the short-circuit circuit using the second bypass valve.

According to a further feature of the invention, the hydraulic drive has pressure sensors that detect a change in a load condition, so that the swivel angle of the hydraulic machine can be adjusted.

The change in the load condition is detected by the pressure sensors. This happens, for example, when the boom touches the ground. In this case, the hydraulic machine is in recuperative mode and pivoted to the negative. If the soil is to be tilled, an active load on the cylinder by the hydraulic machine is required. Then the swivel angle of the hydraulic machine is adjusted into the positive range. In addition, the main control slide is pushed into the further working position and the two bypass valves are closed. This means that the hydraulic drive is in "classic lowering" mode. As a result, the hydraulic machine can quickly apply pressure and extend the consumer cylinders and thus actively actuate the consumer.

According to a further feature of the invention, the hydraulic drive has the pressure relief valves for releasing pressure peaks, in which a load pressure of the relief line acts in the closing direction on a respective end face of the pressure relief valves.

The pressure relief valve is connected to the pump channel via the relief line. During recuperative lowering, the full load pressure is present in the pump channel. This increases the pressure in the relief line. If the load pressure were applied to the end face in the opening direction of the pressure relief valve, the pressure relief valve would open when the load pressure is higher than the set compressive force equivalent of the spring of the pressure relief valve. This would result in an uncontrolled and uncontrollable lowering behavior of the boom have as a consequence. Since the load pressure of the pump channel acts on the face in the closing direction, the pressure relief valve remains in the closed position at the full load pressure of the pump channel and no flow can flow uncontrolled from the boom into the return channel. As a result, the speed of the lowering can be controlled and controlled by the hydraulic machine.

In other words, the load pressure in the pump channel during lowering is reported in a load pressure reporting channel via the hydraulic machine capable of mooring/pivoting. This prevents the pressure relief valve from opening.

The pressure spikes released by the pressure relief valve result from the fact that movement of the swing angle of the hydraulic machine is slower than movement when closing valves.

According to a further feature of the invention, the hydraulic pump and the hydraulic machine are arranged on a common drive shaft. As a result, only one electric motor, which drives the two hydraulic pumps, and one inverter are required. This saves costs and components.

According to a further feature of the invention, the hydraulic pump and the hydraulic machine are each arranged on different drive shafts. As a result, an optimal speed can be achieved for each individual control circuit, which cannot be achieved in every operating situation with a common drive shaft.

According to a further feature of the invention, the hydraulic pump is a variable displacement pump. This ensures a required volume flow in the second control circuit.

According to a further feature of the invention, the hydraulic pump is a fixed displacement pump. This has the advantage that the hydraulic pump always works with its maximum displacement. As a result, the hydraulic pump is operated efficiently and has cost advantages over an adjustable pump.

According to a further feature of the invention, the hydraulic machine is adjustable. This ensures a required volume flow in the first control circuit.

According to a further feature of the invention, the hydraulic machine capable of pivoting is constant. As a result, the hydraulic machine always works with its maximum displacement. As a result, the hydraulic machine is operated efficiently and has cost advantages over an adjustable pump.

According to a further feature of the invention, the pumps for the secondary consumers are arranged on a common drive shaft with the hydraulic machine and the hydraulic pump. This saves costs because no additional drive shaft, electric motor and inverter are required.

According to a further feature of the invention, the pumps for the secondary consumers are arranged on a different drive shaft than the hydraulic pump and the hydraulic machine. This means that a lower voltage can be realized with an additional converter. This means that cheaper electric motors can be used. In this way, the pumps for the auxiliary consumers do not have to be operated constantly when the hydraulic pump and the hydraulic machine are driven.

According to a further feature of the invention, the hydraulic pump is arranged on a drive shaft which is hydraulically driven to enable regeneration of energy. As a result, the energetic efficiency can be increased.

According to a further feature of the invention, the hydraulic drive has a hydraulic slewing gear. This enables the slewing gear to regenerate. As a result, the hydraulic slewing gear is almost as energy-efficient as an electric slewing gear. A hydraulic slewing gear has a cost advantage over an electric slewing gear.

According to a further feature of the invention, the hydraulic slewing gear is designed in a closed circuit. In the closed circuit, the hydraulic drive has an additional hydraulic pump that drives a hydraulic machine of the slewing gear. The hydraulic machine can be designed with a constant or a variable displacement.

According to a further feature of the invention, the hydraulic slewing gear is designed in an open circuit. The system is regulated secondarily. The additional hydraulic pump ensures a volume flow to the hydraulic machine of the slewing gear. The displacement of the hydraulic machine is controlled by a speed controller. It is also conceivable to dispense with the additional hydraulic pump. In this case, the hydraulic pump that drives the slewing gear must be able to pivot. The hydraulic machine of the slewing gear can be operated with secondary control.

character list

• 1 shows a circuit diagram of a hydraulic drive according to the disclosure according to a first embodiment;
• 2 shows a schematic representation of selected components of the disclosure according to the hydraulic drive according to the first embodiment;
• 3 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a second embodiment;
• 4 shows a schematic representation of selected components of the disclosed hydraulic drive according to a third embodiment;
• 5 shows a schematic representation of selected components of the disclosed hydraulic drive according to a fourth embodiment;
• 6 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a fifth embodiment;
• 7 shows a schematic representation of selected components of the disclosed hydraulic drive according to a sixth embodiment;
• 8th shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a seventh embodiment;
• 9 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to an eighth embodiment; and
• 10 shows a schematic representation of selected components of the disclosed hydraulic drive according to a ninth embodiment.

Detailed description of the embodiments

Hereinafter, the embodiments of the present invention will be described based on the accompanying figures.

First embodiment

1 shows a circuit diagram of a hydraulic drive 1 according to the disclosure. The hydraulic drive 1 is used in a (mobile) electrified work machine (not shown), preferably an excavator.

The hydraulic drive 1 has a hydraulic pump 2 . This can be either a variable displacement pump or a fixed displacement pump. The hydraulic drive 1 also has a hydraulic machine 4 that can pivot. The hydraulic pump 2 and the hydraulic machine 4 are both arranged on a common drive shaft 6 which is driven by an electric motor 8 . An inverter 10 is connected upstream of the electric motor 8 . In this case, the inverter 10 is supplied by a rechargeable battery 12 via an electrical intermediate circuit. The battery 12 also supplies an electric rotary drive 14. Additional pumps 16, 18 for secondary consumers are also arranged on the drive shaft 6. The secondary consumers are, for example, a line for providing the control pressure or a steering system and/or a brake.

The hydraulic machine 4 is connected to a pump line 20 . The pump line 20 is in turn connected to a main control slide 22 . The main control slide 22 is designed as an 8/3-way proportional control valve. A port of the main spool 22 is connected to a port A. On the outlet side, the main control slide 22 is connected to a line 24 to an individual pressure compensator 26 . The main control slide 22 has a connection for a line 28 from the individual pressure compensator 26 back to the main control slide 22. In a spring-biased basic position, all connections of the main control slide 22 are blocked. A connection from the pump line 20 to the line 24 to the individual pressure compensator 26 and a connection from line 28 from the individual pressure compensator 26 to the connection A are interrupted in the basic position. In an electrically adjustably actuated working position of the main control slide 22, the pump line 20 is connected to the line 24 to the individual pressure compensator 26 in a hydraulically throttled manner. The line 28 from the individual pressure compensator 26 to the main control slide 22 is hydraulically connected to the port A. A port B is hydraulically connected to a return passage 30 . Secondary pressure relief valves 31 are positioned between ports A and B and main control slide 22, via which a setpoint pressure value can be set on the lines.

In a further electrically adjustable actuated working position of the main control slide 22, which represents an active lowering, the pump line 20 is connected to the line 24 to the individual pressure compensator 26 in a hydraulically throttled manner. Line 28 from individual pressure compensator 26 to main spool valve 22 is hydraulically connected to port B. The connection A is hydraulically connected to the return channel 30 . A connection a relief line or a load pressure reporting channel 32 is hydraulically connected to the line to port B via a check valve.

The individual pressure compensator 26 is designed as a 3/3-way proportional control valve. The individual pressure compensator 26 is connected to the main slide 22 via the line 24 on an input side. A connection to the relief line 32 and the line 28 back to the main slide 22 are arranged on an output side of the individual pressure compensator 26 . In a spring-loaded basic position, the hydraulic connections are blocked. The individual pressure compensator 26 is pushed hydraulically into a working position by a pressure in the line 24 between the main control slide 22 and the individual pressure compensator 26 . In the working position of the individual pressure compensator 26, the line 24 is hydraulically connected to the line 28 back to the main slide 22. The connection of the line 24 from the main slide 22 is connected to the relief line 32 in a hydraulically throttled manner. In the line 28 between the individual pressure compensator 26 and the main control slide 22, two check valves 34, 36 are arranged.

If the line 24 is pressurized because the main control slide 22 is in the working position, the individual pressure compensator 26 is actuated hydraulically. As a result, the lines 24, 28 are hydraulically connected from the main control slide 22 via the individual pressure compensator 26 back to the main slide 22. The port A is thus hydraulically connected to the pump line 20 via the main control slide 22 and the individual pressure compensator 26 .

Port A is connected to base sides 42 of consumer cylinders 44 via two pipe rupture safety valves 38 and 40 (represented in simplified form). The piston sides 45 of the consumer cylinders 44 are connected to port B. Port B is connected to a port of main control slide 22 . In the working position of the main control slide 22, the port B is hydraulically connected to the return channel 30.

The bottom sides 42 of the consumer cylinders are connected to a first bypass valve 46 . The first bypass valve 46 is designed as a 2/2-way proportional control valve. In a spring-loaded basic position, the connections of the first bypass valve 46 are blocked. In an electrically actuated and adjustable or piloted working position, the bottom sides 42 of the consumer cylinders 44 are connected to a pump channel 48 . The pump channel 48 is connected to the hydraulic machine 4 . A check valve 50 is arranged in the pump channel 48 so that no fluid flow from the hydraulic machine 4 into the pump channel 48 is possible.

The hydraulic drive 1 has a second bypass valve 52, which is also designed as a 2/2-way proportional control valve. The second bypass valve 52 is arranged between the base side 42 and the piston side 45 of the consumer cylinder 44 . In a basic position, the second bypass valve 52 blocks. In a working position, the second bypass valve 52 connects the piston side 45 with the bottom side 42 of the consumer cylinder 44.

Pressure sensors 54 are arranged in the lines from the connections A and B to the base side 42 or the piston side 45 of the consumer cylinder 44 .

The load-sensing line 32 with a check valve 56 branches off from the pump channel 48 . The load reporting line 32 is connected to the return channel 30 via a pressure relief valve (load pressure reporting valve) 58 . A throttle 60 is arranged between the check valve 56 and the pressure-limiting valve 58 . A line with a throttle 63 also branches off from the pump channel 48 . The line leads to a tank 74 and has the function of a relief orifice.

The hydraulic connection from the hydraulic machine to the consumer cylinders 44 corresponds to a first control circuit 5.

The hydraulic pump 2 is connected to a summing valve 62 via a second pump line 61 . The summing valve 62 can be designed as a 3/3-way proportional control valve. On the outlet side, the summing valve 62 has a connection to the pump line 20 . In a basic position, the summing valve 62 blocks. In an electrically actuated working position of the summing valve 62, the second pump line 61 is hydraulically connected to the first pump line 20. A line leads from the line from the summation valve 62 to the first pump line 20 to a pressure relief valve or unloading valve 64. The pressure relief valve 64 is designed as a 2/2-way proportional control valve and is connected to the return channel 30 on the output side. The pressure relief valve 64 essentially has the function of a pressure relief valve. In a basic position, the pressure relief valve 64 blocks the connection between the line from the summation valve 62 to the pump line 20 and the return channel 30. In a hydraulically actuated working position of the pressure relief valve 64, the line from the summation valve 62 to the pump line 20 is hydraulically connected to the return channel 30 . This relieves the pressure relief valve 64 the line from the summing valve 62 at too high a pressure.

The hydraulic drive 1 has a hydraulic accumulator 66 . The hydraulic accumulator 66 is connected to the pump channel 48 and thus to the hydraulic machine 4 . The line between the hydraulic accumulator 66 and the hydraulic machine 4 has a check valve 67 and a pressure sensor 68 . The hydraulic accumulator 66 is connected to the return duct 30 via a check valve 70 and can be filled from this. Pressure relief valves 72, 73 are used to preload the return channel 30. In the line to the hydraulic accumulator 66, a pressure sensor 71 is attached. A return volume flow is emptied into a tank 74 via the pressure-limiting valve 73 via a filter 75 .

The hydraulic drive 1 has a retarder valve section 76 . The retarder valve section 76 has a main control slide 78 and an individual pressure compensator 80 which are structurally identical to the main control slide 22 and the individual pressure compensator 26 . The main control spool 78 is connected to an A port. A load pressure is specified via a pressure-limiting valve 82 between port A and the return duct 30 . As a result, additional hydraulic power is required, which is provided by the hydraulic pump 2 . Thus, energy that was regenerated and transferred to the hydraulic pump 2 by the retraction of the consumer cylinders 44 can be dissipated through the retarder valve section 76 .

The hydraulic drive 1 has a further valve section 84 for the further consumer. The further valve section 84 is the second control circuit. The structure of the further valve section 84 is essentially identical to the structure of the first control circuit. The hydraulic pump 2 builds up hydraulic pressure via the second pump line 61 . The hydraulic pump 2 is connected to a port A by a main control slide 86 and an individual pressure compensator 88 . This is connected to a bottom side of another consumer cylinder 90 . A piston side of the further consumer cylinder 90 is connected to a connection B. This is connected to the main control slide 86. Pressure relief valves 92 are positioned between the main spool valve 86 and ports A and B. A relief channel 94 leads to the return channel 30 via a pressure relief valve 96 . The further valve section 84 also has a pressure relief valve 98 .

• Durch die Hydraulikmaschine 4 wird ein Druck in der Pumpenleitung 20 beaufschlagt. Wenn der Hauptsteuerschieber 22 durch Betätigung eines elektrischen Aktors in seine Arbeitsstellung geschoben wird, wird die Leitung 24 von dem Hauptsteuerschieber 22 zu der Individualdruckwaage 26 hydraulisch gedrosselt mit (Druck beaufschlagt) Volumenstrom versorgt. Dadurch wird die Individualdruckwaage 22 in ihre Arbeitsstellung geschoben und die Leitung 28 von der Individualdruckwaage 26 zurück zu dem Hauptsteuerschieber 22 ist hydraulisch mit der Pumpenleitung 20 verbunden. Die Bodenseiten 42 der Verbraucherzylinder 44 werden so über den Hauptsteuerschieber 22 und die Individualdruckwaage 26 mit (Druck beaufschlagt) Volumenstrom versorgt. Die Kolbenseiten 45 der Verbraucherzylinder 44 sind über den Anschluss B mit dem Rücklaufkanal 30 verbunden. Damit wird ein Rücklauf von Arbeitsfluid von den Kolbenseiten 45 in den Rücklaufkanal 30 ermöglicht.

The arrangement described above works as follows:

• A pressure is applied in the pump line 20 by the hydraulic machine 4 . When the main control slide 22 is pushed into its working position by actuating an electric actuator, the line 24 from the main control slide 22 to the individual pressure compensator 26 is supplied with a hydraulically throttled (pressurized) volume flow. As a result, the individual pressure compensator 22 is pushed into its working position and the line 28 from the individual pressure compensator 26 back to the main control slide 22 is hydraulically connected to the pump line 20 . The bottom sides 42 of the consumer cylinders 44 are thus supplied with (pressurized) volume flow via the main control slide 22 and the individual pressure compensator 26 . The piston sides 45 of the consumer cylinders 44 are connected to the return channel 30 via the connection B. This allows working fluid to flow back from the piston sides 45 into the return channel 30 .

Through this process, the consumer or boom is raised in the classic manner or the consumer cylinder is extended. This mode of operation corresponds to that of a known LUDV valve from the prior art.

If energy recovery is not possible when lowering the boom, for example if the charge level of the battery 12 is too high, the "classic" lowering can take place, i.e. via the main control slide 22. Here, the main control slide 22 is pushed into the other working position. As a result, the pump line 20 is hydraulically connected to the port B via the individual pressure compensator 26 . In the process, a (pressure) volume flow is applied to the piston sides 45 of the consumer cylinders 44 via the hydraulic machine 4 . The bottom sides 42 of the consumer cylinders 44 are hydraulically connected to the return channel 30 via the main control slide 22 .

With active, ie driving loads, the return orifice of the main control slide 22 (in conjunction with the pipe rupture safety valves 38 and 40) specifies the lowering speed of the boom. The lowering speed of the jib is the retracting speed of the consumer cylinders 44 and the hydraulic machine 4 generally adjusts a significantly smaller inflow volume flow adjusted for the cylinder area ratio. The missing differential volume flow can be obtained from the return volume flow by the regeneration circuit located in the main control slide 22 . For this purpose, the relief line 32 is hydraulically connected to the line to port B via the main control slide 22 .

The pump volume flow can be lowered by the regeneration circuit from the relief line 32 and the energetic efficiency can be improved above all in a possible summation operation. If a load change occurs, for example, after the shovel (not shown) comes into contact with the ground to be worked, the retraction speed of the boom is subsequently specified via the inflow volume flow provided by the hydraulic machine 4 . The operator accepts the slowing down of the boom speed that occurs as a result, since the digging or jacking process usually begins at this point and increased precision is required.

The boom can be lowered recuperatively. The pipe rupture safety valves 38, 40 and the first bypass valve 46 are shifted by electric actuators. As a result, the bottom sides 42 of the consumer cylinders 44 are hydraulically connected to the pump channel 48 . The hydraulic machine 4 swings through and functions as a hydraulic motor. This regenerates energy. For the recuperative lowering of the boom, the ability to swivel through the swivel mechanism of the hydraulic machine 4 is used, as a result of which it can be operated as a hydraulic motor without the direction of rotation of the drive shaft driving it having to be reversed.

Design and system-related leakage points occur in the hydraulic drive. These occur, for example, in the relief line 32, in the first bypass valve 46 between the base sides 42 of the consumer cylinder 46 and the pump channel 48 and in the hydraulic machine 4. The pressure in the pump channel 48 can be reduced by these leaks, which leads to an abrupt lowering of the boom. This contradicts fine controllability of the boom and is unacceptable. The influence of leakage is compensated by the pipe rupture safety valves 38 and 40 .

In other words, due to the design and system-related leakage points present in the displacement unit, a volume flow offset occurs as a result of the leakage volume flow that occurs as soon as pressure is present in the pump channel 48 . This circumstance would result in a jump in speed build-up, which would contradict fine controllability and would therefore be unacceptable. In order to compensate for the influence of leakage, the pipe rupture safety valves 38 and 40 are used.

The pipe rupture safety valves 38 and 40 are designed as proportional valves and can be actuated either directly electrically or in a piloted manner, for example as a valvistor, or can be supplied with an external control pressure via a line. Via the fine control range of the pipe rupture safety valves 38 and 40, an opening cross section is then set very sensitively, which directly influences the lowering speed of the consumer cylinders 44, in that a volume flow flows from the bottom sides 42 of the consumer cylinders 44 into the parallel pump channel 48. For this purpose, the first bypass valve 46 has to be opened, which connects the bottom sides 42 of the consumer cylinders 44 to the pump channel 48 . The first bypass valve 46 can be designed as a proportional valve, so that it can theoretically also represent the functionality of leakage compensation in excavators without the need to use pipe rupture safety valves (for example in excavators in Japan).

The first bypass valve 46 is necessary to block the pump channel 48 via the valve unit during the classic lowering process and thus to counteract an increase in pump pressure caused by the necessary connection of the pump channel 48 to the relief line 32 and to prevent uncontrolled lowering. If the first bypass valve 46 is fully open, a pressure can build up in the pump channel 48 as a result of the volume flow flowing out of the pipe rupture safety valves 38 and 40 if the system and component-related leaks have been compensated for. The leaks caused by the system and components appear via the hydraulic machine 4, the check valve 56 and the throttle 63, while the swivel angle of the hydraulic machine 4 is swiveled back to zero in this phase of speed introduction. The pressure in the pump channel 48 drops as a result of the leaks. When the first bypass valve 46 is fully open, a pressure loss in the pump channel 48 leads to an uncontrolled lowering of the consumer. From a certain pressure to be defined, the swivel angle of the hydraulic machine 4 is swiveled out in the direction of motor operation as the speed specified by the operator continues to increase, and the pipe rupture safety valves 38 and 40 increasingly begin to set a large opening cross section. In a third and final phase or a specific speed preset by the operator to be adapted beforehand to the circumstances of the system, only the swivel angle of the hydraulic machine 4 still determines the speed of the boom. The pipe rupture safety valves 38 and 40 are fully open here and release a maximum opening cross section, so that only very small pressure losses occur across them during the recuperative lowering.

Since the hydraulic machine 4 acts as a hydraulic motor and thus as a working fluid flow sink during the recuperative lowering, it cannot be used as a volume flow source for filling the piston sides 45 of the consumer cylinders 44 . A summation with the working fluid flow from the hydraulic pump 2 via the summation valve 62 also makes little sense for energy reasons, since especially during the parallel operation of several consumers due to the low pressure of the volume flow required by the boom, there is a high power loss at the same time higher pressures of the parallel consumers. For this reason, the hydraulic drive 1 has the second proportionally adjustable bypass or regeneration valve 52 which connects the bottom side 42 to the piston side 45 of the consumer cylinder 44 . As a result, the two cylinder chambers 42 and 45 can be short-circuited (if the opening cross section of the second bypass valve 52 is large enough), as a result of which the pressures in the two cylinder chambers 42 and 45 adjust and the pressure increases due to a virtual reduction in the cylinder surface.

Two advantages result from the short-circuit connection between the base sides 42 and the piston sides 45 of the consumer cylinders 44 through the second bypass valve 52: On the one hand, the piston side 45 can be filled without an additional working fluid flow provided by a pump, for example, and on the other hand the volume flow decreases , which must flow back to the hydraulic machine 4 from the bottom sides 42 of the consumer cylinders 44 . This allows the speed to be reduced and the displacement of the hydraulic machine 4 to be set to a maximum value. In addition, the pressure losses in the resistance of the pressure lines caused by the volume flow are reduced. These circumstances contribute to an increase in the overall energy efficiency of the system, essentially due to component-related reductions in mechanical losses (reduction in splashing losses in the hydraulic machine 4, more favorable force or lever ratios on the axial pistons of the hydraulic machine 4, lower bearing friction in the electric machine 8 and the hydraulic machine 4, etc.) at.

Another task of the second bypass valve 52 is to limit the maximum pressure of the pressures that occur in the two cylinder chambers 42 and 45 in the event of a short circuit Short-circuit circuit results in a higher pressure in the cylinder chambers 42 and 45 than the opening pressure set at the secondary pressure relief valves 31, the opening cross-section of the second bypass valve 52 is adjusted by appropriate control so that the resulting flow loss in the pump channel 48 reduces the pressures in the cylinder chambers 42 and 45 effected. The setpoint value set for controlling the pressures must be below the opening pressures set at the secondary pressure relief valves 31 in order to prevent the cylinder from being lowered in an uncontrolled manner. The pressure sensors 53 and 54 are used for the control, with the pressure sensor 54 on the base being absolutely necessary and the pressure sensor 53 on the piston side being included only to improve the control behavior.

A further challenge of this control arrangement lies in switching between recuperative and classic retraction of consumer cylinders 44 when load changes occur. These load changes occur primarily when the boom or the tool, usually the bucket, touches the ground that is to be processed. This changes the pressure conditions in the consumer cylinders 44 and an active load becomes a passive load that has to be driven. However, since the hydraulic machine 4 has pivoted into the negative as a result of the recuperative lowering process, the change in the load state must be detected by the pressure sensors 53, 54 and 68 and recognized by an evaluation unit. In this case, the swivel angle of the hydraulic machine 4 is adjusted into the positive range and the main control slide 22 is adjusted in the direction of lowering the boom, and the two bypass valves 46 and 52 are closed. The boom can now be lowered "classically" and a corresponding force in the Giving direction of movement. Since a jump in speed would also be felt in the case of a complete lowering process in the classic sense due to the system, the delay times for a switchover process from recuperative to classic lowering are also accepted.

Additional measures must be provided in the hydraulic drive 1 in order to avoid uncontrolled lowering. The reason for this is the use of the pressure relief valves 64 and 98, which cut off or prevent pressure peaks when the volumetric flow specifications change and take over a pressure limitation of the line from the summation valve 62. The pressure relief valves 64 and 98 thus take over a primary pressure limitation as a kind of pilot-operated pressure relief valve in conjunction with the load pressure reporting pressure relief valve 58. The pressure peaks result from a compared to the closing of valves slower movement of the pump swivel angle when swiveling back. Since the main control slide 22 is not moved during recuperative lowering and therefore no control edges are opened either, no load pressure can form in the relief line 32 . However, during the recuperative lowering, the pressure in the pump channel 48 increases to the value of the load pressure after the pipe rupture safety valves 38 and 40 are fully open and the first bypass valve 46 is open. This increases the pressure on the opening face of the pressure relief valve 64, which can thereby open fully as soon as the load pressure is higher than the set compressive force equivalent of the spring of the pressure relief valve 64. This would result in an uncontrolled and uncontrollable lowering behavior of the boom, which cannot be tolerated for safety reasons alone.

For this reason, an interconnection is proposed in the present disclosure, which consists essentially of a connection of the pump channel 48 with the relief line 32 of the boom and the check valves 50 and 56 and the throttle 63 . The aim is that the load pressure applied to the hydraulic machine 4 during the recuperative lowering is also reported to the common relief line 32 and the summation valve 62 and is also applied to the end face of the pressure relief valve 64 in the closing direction. Since the prestressing force of the spring also acts in this direction, the pressure relief valve 64 remains in a closed state and no volume flow can flow uncontrolled from the boom into the return channel 30 and the speed can still be controlled by the hydraulic machine 4 .

The check valve 56 prevents a backflow of working fluid from the load-sensing line 32 into the pump channel 48, where it would otherwise flow via the throttle 63 into the return 30 and thus no pressure could build up in the load-sensing line 32 during normal operation. The throttle 63 serves to avoid possible pressure restraints in the pump channel 48 and as a damping orifice during the pressure build-up during recuperative lowering of the boom. The check valve 50 prevents the working fluid from flowing into the pump channel 48 during normal operation of the assembly (e.g. when raising the boom) and thus raising the pump pressure to maximum pressure should the system be implemented as an eLS system.

Another special feature of this control arrangement is the wiring and arrangement of components to protect the hydraulic machine 4. This is necessary because changes in load and movement can cause conditions that can lead to cavitation at the pressure inlet of the hydraulic machine 4 and thus damage it because this is not designed for this. These cavitation states can occur when, for example, a rapid reversal of movement is to take place from rapid lifting to rapid lowering of the boom. Since the boom is very sluggish, it follows its activation with a time delay and so when the hydraulic machine 4 pivots, working fluid can already be sucked out of the pump channel 48 or the pump line 20 before the consumer cylinders 44 have changed their direction of movement and can fill the pump channel 48 with working fluid .

This effect also occurs when the load conditions change, as has already been explained in the description of the switchover from recuperative to conventional retraction of consumer cylinders 44 . Here, too, the hydraulic machine 4 can suck working fluid out of the pump channel 48 or the pump line 20 when the tool touches the ground and the boom can therefore not be lowered any further. The two changes of state described here can be detected in part with the pressure sensors 53, 54 and 68, but due to the different dynamics of the boom, pressure build-up and swivel angle adjustment, there is always the risk of cavitation, especially in the transient behavior. In order to protect the hydraulic machine 4 in these transient areas in particular and to prevent cavitation, working fluid can be taken from the hydraulic accumulator 66 via the check valve 67 and pumped into the pump line 20 of the hydraulic machine 4 . This happens automatically as soon as the pressure in the pump line 20 is lower than the pressure in the hydraulic accumulator 66.

In order to fill the hydraulic accumulator 66, the target values for the opening pressures of the two pressure-limiting valves or preloaded check valves 72 and 73 are increased to a value that allows the return volume flow from the return channel 30 through the check valve 70 to fill the hydraulic accumulator 66 until the maximum specified accumulator pressure has been reached. The pressure limiting valves 72 and 73 are used to preload the return channel 30, with the pressure limiting valve 73 being set to a lower value than the pressure limiting valve 72, so that a large part of the return volume flow can flow through the filter 75 on the way to the tank 74. If the flow resistance across the filter 58 increases too much, the pressure relief valve 72 opens and allows the excess volume flow to protect the filter 58 unfiltered into the tank 74. The actual accumulator pressure can be taken from the pressure sensor 71 become men. If the accumulator 45 is filled with the required amount of working fluid, the target values for the pressure-limiting valves 72 and 73 can be reset to their standard values. The actuation of the pressure-limiting valves 72 and 73 can be realized with an externally applied control pressure or directly electrically actuated. The working fluid flow for filling the hydraulic accumulator 66 can take place directly from the volume flows flowing back from the consumers into the return channel 30 or, if these are insufficient or not available, through the hydraulic pump 2 and/or the hydraulic machine 4.

In order to achieve this, the hydraulic pump must be overridden beyond the actual total volume flow requirement of all actuated consumers, which can be realized very advantageously with so-called flow matching systems. The volume flow not required by the consumers can then flow into the return channel 30 via the respective pressure relief valves 64 and/or 98 . A further possibility is the actuation of the retarder valve section 76, in which the main control slide 78 is actuated in direction A and the volume flow can flow into the return channel 30 via the non-actuated pressure-limiting valve 82. The two connections A and B of the retarder valve section 76 are closed. Before the boom can be lowered recuperatively, the hydraulic accumulator 66 must always be full. This can be checked with the pressure sensor 71.

First and foremost, the retarder function should fulfill the functionality of being able to dissipate energy at the moment when energy can be recovered electrically from states of motion or their changes, but cannot be temporarily stored in the battery. This is the case, for example, when the battery/rechargeable battery 12 is fully charged or when the battery is defective. This is the case, for example, when the rotating movement of the upper carriage is decelerated and the resulting energy cannot be fed into the electrical storage device/battery 12. Or even if the electrical energy can no longer be absorbed during the recuperative lowering, the energy can be destroyed accordingly by the retarder. A volume flow can be measured by the main control slide 78 in the retarder valve section 76 via the inlet orifice in the direction of the connection A and a load pressure can be specified by specifying the opening pressure of the pressure-limiting valve 82 . This causes additional hydraulic power, which has to be provided by the hydraulic pump 2 . It is thus possible to transmit the energy or power recovered by the recuperative lowering of the boom from the hydraulic machine 4 directly to the hydraulic pump 2, which is intended to be dissipated by the retarder section. The situation is similar with the power fed back from the slewing gear. This is picked up by the electric motor 8 and also forwarded to the hydraulic pump 2 through the connected drive shaft 6 .

2 shows a schematic representation of selected components of the disclosed hydraulic drive according to the first embodiment. The hydraulic pump 2 is a variable displacement pump and the hydraulic machine 4 is a variable hydraulic machine. The hydraulic pump 2 and the hydraulic machine 4 are arranged on the same drive shaft 6 . The drive shaft 6 is driven by the electric motor 8, which is preceded by an inverter 10. The fluid flows of the hydraulic pump 2 and the hydraulic machine 4 are added by the summation valve 62 . The pumps 16, 18 for the secondary consumers are also arranged on the common drive shaft 6.

Second embodiment

In the following embodiments, components that are identical to those in the first embodiment will not be described in detail. If the functional relationships between the individual components are the same in the embodiments, only the differences between the individual embodiments will be discussed.

3 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a second embodiment. In the second embodiment, the hydraulic pump 2 is a fixed displacement pump. The arrangement of the individual components and the functional relationships between the components is the same in the second embodiment as in the first embodiment.

The volume flow requirement is achieved in this circuit by a speed-variable control of the electric motor 8, so that in many cases or operating states during use of the excavator, a needs-based supply can be achieved. The advantage lies in more efficient operation since the hydraulic pump 2 always operates at its maximum displacement and also in lower costs since hydraulic pumps with a constant displacement have a cost advantage over hydraulic pumps with a variable displacement.

Third embodiment

4 shows a schematic representation of selected components of the disclosed hydraulic drive according to a third embodiment. The hydraulic machine 4 and the pumps 16, 18 for the secondary consumers are arranged on the same drive shaft 6. The hydraulic pump 2, which is a variable displacement pump, is arranged on its own drive shaft 100. The hydraulic pump 2 is driven by its own electric motor 102 with its own inverter 104 .

A better adjustment of an optimal speed for each individual control circuit can be achieved in this way, which is not achieved in every operating situation by a common drive shaft. A summation of the volume flows within the summation valve 62 is also possible, so that the hydraulic pumps 2, 4 and above all the electric motors 8, 102 can be smaller in their dimensions than in the first embodiment. The pumps 16 and 18 for providing the volume flows for the control pressure supply and for the steering/brake remain on a common drive shaft with one of the hydraulic pumps. In the case shown, this is the hydraulic machine 4 for the first consumer, ie the boom. However, it would also be possible to connect the pumps 16 and 18 to the drive shaft 100 of the hydraulic pump 4 for the other consumers, ie the stick and bucket.

Fourth embodiment

5 shows a schematic representation of selected components of the disclosed hydraulic drive according to a fourth embodiment. The fourth embodiment is similar to the third embodiment except that the hydraulic pump 2 is a fixed displacement pump. The hydraulic pump 2 thus has a constant displacement.

Fifth embodiment

6 shows a schematic representation of selected components of the disclosed hydraulic drive according to a fifth embodiment. In the fifth embodiment, the hydraulic machine 4 is not adjustable. The hydraulic machine 4 is driven by the electric motor 8 . The hydraulic pump 2 is also not adjustable and is driven by the additional electric motor 102 . The pumps 16, 18 for the secondary consumers are driven by yet another electric motor 106 with an inverter 108. The hydraulic pumps 2, 4 and 16 and 18 are therefore each arranged on different drive shafts.

In this variant, the control pressure pump 16 and the pump 18 for supplying the steering and brakes are driven by the separate electric motor 106 . However, the electric motor 106 does not necessarily have to be operated with the voltage of the intermediate circuit. An additional DC/DC converter (not shown) can also be used to supply the inverter 108 with a significantly lower voltage and thus cheaper electric motors from an expanded product portfolio (for example 12V, 24V or 48V) can also be used. The advantage of separating the supply of drive power for control pressure and steering/brakes is that pumps 16 and 18 are no longer constantly driven when hydraulic pumps 2 and 4 have to be driven. As a result, power losses can be avoided, which mainly result from neutral circulation current losses in the hydraulic drive. A further advantageous addition can be seen in the integration of a hydraulic accumulator 110 in the control pressure line. As a result, corresponding hydraulic energy can be stored and the hydraulic motor or pump 16 can only be operated when the pressure in the hydraulic accumulator 110 drops below a previously specified minimum value. Only then is the electric motor 106 switched on and fills up the hydraulic accumulator 110 again. The advantage here is that the pumps 16 and 18 and the electric motor 106 are not constantly in operation and while they are in operation, however, they can be operated in an optimal working range. In a further embodiment of the system, the hydraulic pumps 2 and 4 can also be implemented with variable swivel angles.

Sixth embodiment

7 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a sixth embodiment. The arrangement of the components in the sixth embodiment is the same as the arrangement of the components in the fifth embodiment. However, in the sixth embodiment, the hydraulic machine 4 and the hydraulic pump 2 are variable. In addition, the hydraulic machine 4 and the hydraulic pump 2 can be connected to one another via a clutch 12 . This results in the possibility of direct mechanical recuperation when lowering the boom in the direction of the hydraulic pump 2, which can improve energy efficiency if the operating situation allows it. In addition, the individual electrical machines 8 and 102 and inverters 10 and 104 do not have to be at maximum capacity tion / maximum torque of the hydraulic consumers are designed, but can be interconnected if necessary, if one consumer or more consumers request a higher moment, for example, than the maximum torque for which the respective electric machines are designed. This means that smaller and, above all, cheaper machines and inverters can be used.

Seventh embodiment

8th shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to a seventh embodiment. The previously listed embodiments all have a slewing gear or a hydraulic pump, which is driven directly electrically. For cost reasons, however, it may make sense to run the slewing gear hydraulically if it is also capable of regeneration and can therefore be operated with almost the same energy efficiency as a purely electric slewing gear. 8th shows a variant in which the slewing gear is designed in a closed circuit. Here, an additional hydraulic pump 114 is integrated on the drive shaft 100 of the stick/bucket drive unit and drives a hydrostatic pump 112, which can be designed either with constant or variable displacement. The hydraulic pump 114 must be designed with an adjustable displacement. The arrangement of the pumps 16 and 18 for the secondary consumers is identical to that in the fifth and sixth embodiment.

Eighth embodiment

9 shows a schematic representation of selected components of the hydraulic drive according to the disclosure according to an eighth embodiment. The eighth embodiment is an expression of the open loop system as a secondary controlled system. Here, the hydrostatic hydraulic pump 114 (with variable displacement) provides a pressure in the supply line to the hydrostatic pump 112, which is controlled by, for example, a speed control in its displacement and can realize a movement by torque build-up.

Ninth embodiment

10 shows a schematic representation of selected components of the disclosed hydraulic drive according to a ninth embodiment. It is also conceivable to dispense with the hydrostatic hydraulic pump 114 of the eighth embodiment and, if necessary, to provide a required pressure in the common control circuit for the stick/bucket circuit and the slewing gear by means of the hydraulic pump 2 (this must then be designed to be swivellable like the hydraulic machine 4). build up. The hydrostatic pump 112 can then be operated again with secondary control. Here, for example, the volume flow flowing back when the slewing gear is decelerated can be made available directly to the users, the stick and bucket. In addition, here too the hydraulic machine 4 and the hydraulic pump 2 can be connected to one another via a clutch 12 . This results in the possibility of direct mechanical recuperation when lowering the boom in the direction of the hydraulic pump 2, which can improve energy efficiency if the operating situation allows it. In addition, the individual electrical machines 8 and 102 and inverters 10 and 104 do not have to be designed for the maximum power/maximum torque of the hydraulic consumers, but can be interconnected if necessary, if one consumer or several consumers request a higher torque than the maximum torque, for example which the respective electrical machines are designed. This means that smaller and, above all, cheaper machines and inverters can be used.

The hydraulic drive 1 can be designed as an LUDV system (load-pressure-independent flow valve), but also theoretically as any other hydraulic system common for excavators, e.g. NC, ePC, or for any other pump strategy, e.g. VBO.

The control arrangement described in this disclosure is implemented as a LUDV system and can be operated with an electronic eLS system controller as well as an EFM system. In principle, of course, there is also the possibility of designing this control arrangement as any other throttle-controlled system conceivable when used in excavators. In addition, the pump volume flows of one circuit or one pump can be conveyed or added to the other circuit.

If the system is designed as an EFM system, the inflow orifice of the boom circuit can theoretically be opened completely and the volumetric flow can be set purely via the pump using displacement control. This would have the advantage that a correspondingly low pressure drop can be realized across the inlet orifice and thus the energetic efficiency in these states is improved.

QUOTES INCLUDED IN DESCRIPTION

This list of documents cited by the applicant was generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Patent Literature Cited

• DE 102014216031 A1 [0002]
• EP 3358201 A1 [0002]
• EP 2738397 B1 [0003]

Claims (15)

Hydraulic drive (1) for a mobile working machine, preferably an excavator, comprising: - a hydraulic machine (4) capable of pivoting, which is hydraulically connected to first consumer cylinders (44) and is intended and designed to drive them and use potential energy when retracting or retracting the first consumer cylinder (44), - a hydraulic pump (2), which is hydraulically connected to further consumer cylinders (90) and drives them, characterized in that the hydraulic machine (4) capable of pivoting through a first control circuit (5th ) is connected to the first consumer cylinders (44) and the hydraulic pump (2) is connected to the other consumer cylinders (90) via a second hydraulic control circuit (84) and the control circuits (5, 84) are hydraulically separated from one another. Hydraulic drive (1) after claim 1 , characterized by a summation valve (62) which is provided and adapted to add a fluid flow of the hydraulic pump (2) to a fluid flow of the first hydraulic machine (4) and vice versa. Hydraulic drive (1) after claim 1 or 2 , characterized in that the pressure level in the individual control circuits is different from one another. Hydraulic drive (1) according to one of Claims 1 until 3 , characterized in that the pivotable hydraulic machine (4) regenerates the recovered energy directly to the hydraulic pump (2). Hydraulic drive (1) according to one of Claims 1 until 3 , characterized in that the pivotable hydraulic machine (4) converts the recovered energy into electrical energy. Hydraulic drive (1) according to one of Claims 1 until 5 , characterized by a low-pressure accumulator (66) that provides a minimum pressure that prevents the pivotable hydraulic machine (4) from running in an area with too low a pressure. Hydraulic drive (1) according to one of Claims 1 until 6 , characterized in that the pivotable hydraulic machine (4) is connected via a pump line (20) with a main control slide (22) and individual pressure compensator (26) to bottom sides (42) of the consumer cylinders (44) and piston sides (45) of the consumer cylinders (44) via a pump channel (48) are connected to the hydraulic machine (4) capable of pivoting and the fluid flow of the hydraulic pump (2) is introduced into the pump line (20) via the summation valve (62). Hydraulic drive (1) according to one of Claims 1 until 7 , characterized by pipe rupture safety valves (38, 40) which set an opening cross-section in an electrically controlled manner, by means of which a volume flow from the bottom sides (42) of the consumer cylinders (44) into the pump channel (48) is controlled. Hydraulic drive (1) according to one of Claims 1 until 8th , characterized by a first bypass valve (46) which blocks and opens a connection between the pump channel (48) and the bottom sides (42) of the consumer cylinders (44). Hydraulic drive (1) according to one of Claims 1 until 9 , characterized by a second bypass valve (52), which realizes a connection between the bottom sides (42) and the piston sides (45) of the consumer cylinder (44). Hydraulic drive (1) according to one of Claims 1 until 10 , characterized in that the hydraulic pump (2) realizes a fluid flow to the piston sides (45) of the consumer cylinder (44) during the recuperative lowering of the consumer cylinder (44) via the summation valve (62). Hydraulic drive (1) according to one of Claims 1 until 11 , characterized by pressure sensors (53, 54) which detect a change in a load condition, so that a swivel angle of the hydraulic machine (4) capable of swiveling through is set in a positive range. Hydraulic drive (1) according to one of Claims 1 until 12 characterized by pressure relief valves (64, 98) in which a load pressure of a relief line (32) acts in a closing direction on a respective end face of the pressure relief valves (64, 98). Hydraulic drive (1) according to one of Claims 1 until 13 , characterized in that the hydraulic pump (2) and the pivotable hydraulic machine (4) are arranged on a common drive shaft (6). Hydraulic drive (1) according to one of Claims 1 until 14 , characterized in that pumps (16, 18) are arranged for secondary consumers on a different drive shaft (107) than the hydraulic pump (2) and the hydraulic machine (4) which can be swiveled through, but which through a mech. Coupling can be connected to each other.

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