Patent Application
20100065135
The invention relates to a controlling device for hydraulic consumers with at least one control valve for controlling a supply line for the respective hydraulic consumer and with a tank return line.
Controlling devices such as these are used in particular as so-called mobile directional control valves for controlling hydraulic consumers, such as, for example, working cylinders and hydraulic motors. Some of these consumers always experience the same direction of force of the external load; other loads change the direction of their force in operation. Thus the lifting cylinder of a fork lift always experiences a force directed downward, whereas the hydraulic motor of a slewing gear during acceleration can experience a compressive load and, upon braking, a pulling load because the inert mass of the slewing gear continues to run in the original drive direction.
If, at this point, pulling loads move the consumer more quickly than corresponds to the volumetric flow amount in the supply line, the inlet pressure drops rapidly down to the cavitation pressure and below. This is to be fundamentally prevented.
To counter this, readily available control devices can be purchased on the market which ensure that the cavitation pressure is always reached for the indicated suction action of pulling loads. In the known solutions fluid is supplied by way of an additional feed system to the respectively endangered pressure line as the supply line. This supply, however, only takes place when the feed pressure which has been applied by way of this feed system is greater than the pressure in the endangered supply line plus the sum of all pressure drops at the installed throttle points from the supply line to the endangered line. An additional pump system is often encountered as an additional feed system in hydrostatic drives. One option, which is more economical in comparison, arises when the fluid backflow to the tank as a pressure chamber is retained in conventional valve controls by way of a so-called tank back pressure valve as the control valve and then the required supply volume is taken from this pressure chamber. The disadvantage in these known solutions is the continuing energy loss resulting from the additionally required pump delivery amount and the set back pressure or working capacity of the hydraulic consumer which essentially has been reduced by the back pressure.
In order to at least partially remedy this, DE 43 42 487 B4 discloses a hydrostatic drive system with a consumer of hydraulic energy which can be supplied on both sides, which is located in an open circuit, and whose two ports are assigned at least one brake valve with a replenishing valve which is dynamically connected to it, the replenishing valve enabling supply of a hydraulic medium from the outlet side to the inlet side of the consumer. In the known solution, it is provided that the replenishing valve in the braking phase in which the brake valve can produce an outlet-side pressure can be preloaded to an increased replenishing pressure by the pressure which is produced on the outlet side in the braking phase. In the normal operating state then fluid hydraulic medium can escape without great resistance by way of the replenishment valve to the tank and in the braking phase the replenishment valve is automatically preloaded to a higher opening pressure so that due to the increased replenishment pressure level, external supply to the hydrostatic driving system can be omitted. In this respect, for the known solution the necessity of providing an additional pump system as an auxiliary pump for maintaining a specific inlet-side pressure level is obviated, as already mentioned; however, the known solution with a control valve in a double piston execution is complex and therefore expensive to produce.
DE 42 43 578 A1 discloses commercial vehicle hydraulics, in particular for a refuse collection vehicle, with at least one hydraulic circuit, to which various actuating elements are connected for performing various functions, such as, for example, opening the rear part, lifting and tipping a dumpster, etc. Furthermore, the known solution has a pump which can be driven by a motor or a secondary output of the commercial vehicle, which is coupled to it, for conveying hydraulic oil into the hydraulic circuit. The pump is designed such that its delivery rate can be controlled at least partially independently of the engine speed. With the known solution it is possible, using a control means, to determine the power demand of the actuating members connected to the hydraulic circuit and to always set the delivery rate of the pump such that the engine speed of the commercial vehicle remains as low as possible and is raised only when the power demand is higher; this helps avoid energy losses.
Furthermore, DE 197 35 482 A1 discloses a hydraulic system with a differential cylinder with a piston rod and piston, which separates the piston rod-side pressure chamber and the pressure chamber which is remote from the piston rod from one another. By means of a directional control valve with two consumer ports the two pressure chambers of the differential cylinder can be connected alternately to a source of hydraulic medium and to a tank. Independently of this directional control valve, by means of a quick operating valve the piston rod-side pressure chamber can be connected to the pressure chamber which is remote from the piston rod of the differential cylinder. In the known solution, when the quick operating valve is actuated in the rest position or the working position of the directional control valve in which the pressure chamber of the differential cylinder which is remote from the piston side is supplied with hydraulic medium from the source of a hydraulic medium, so-called collapse of the load is prevented by a check valve which is located in the connection which can be established by way of the quick operating valve between the two pressure chambers of the differential cylinder and which blocks from the pressure chamber which is remote from the piston rod to the piston rod-side pressure chamber. The known solution thus irrespective of the magnitude of the load which is being moved with the differential cylinder and which counteracts the extension of the piston rod thus allows arbitrary actuation of the quick operating valve without endangering anyone and without the risk of damage to the machine so that at any instant a quick traverse motion is possible.
Proceeding from this prior art, the object of the invention is to further improve the known solutions such that in a reliable, energy saving and economical manner harmful cavitations are reliably prevented in any application. This object is achieved with a controlling device with the features of claim 1 in its entirety.
In that, as specified in the characterizing part of claim 1, the control valve is connected to an additional feed line and is designed as a priority valve such that the supply line acquires preference of fluid supply over the tank return line, a type of sensor circuit is implemented which checks whether, depending on the load situation on the hydraulic consumer, there is any demand for supply flow at all. Only when this demand is “sensed” by the sensor circuit, the tank return line is dammed to a required pressure level and the required inlet pressure in each individual case is maintained such that the cavitation pressure is, in any case, exceeded. This also leads to energy saving effects. The solution according to the invention manages with few components and is thus economical to produce and maintain. The use of additional brake valves, as is shown in the prior art, can therefore be omitted. As a result of the mechanically simple structure, reliable operation for each load state is also ensured. Preferably, the sensor circuit is implemented using a compensator as the control valve.
The control valve for the controlling device according to the invention is designed as a priority valve which, as a so-called tank back pressure valve, gives preference to the indicated supply line over the free tank return line. Preferably a check valve which is located between the supply line and the feed line and which opens in the direction of the supply line prevents inadvertent backflow from the supply line into the feed line.
Another, second feed line can be provided for additional and direct supply of the supply line. Preferably the other feed line in the control valve can be influenced by way of the control edge of the valve piston and can be blocked by the control stroke of the control spring of the control valve such that the connection to the supply line is interrupted. Preferably the other, second feed line begins in a channel of the pressure supply and is determined by way of a defined throttle point in its flow behavior.
The control behavior of the controlling device can be improved by a continuously throttled relief line from the supply line into the free tank return line.
Other advantageous configurations of the controlling device according to the invention are the subject matter of the other dependent claims.
The controlling device according to the invention is detailed below using one embodiment as shown in the drawings.
The figures are schematic and not to scale.
The illustrated working cylinder 10 can be, for example, a component of a machine in the form of a wheel loader or the like for raising or lowering an implement in the form of a conventional lifting mechanism with a blade. The hydraulic motor 12, for example, drives a mechanical slewing gear 16 on the basis of a moment of inertia J. With the respective hydraulic motors 12, for example, hydraulic lifts can be actuated, running gears of machinery such as fork lifts can be driven, and the like. The possible uses both for hydraulic working cylinders and also for hydraulic motors are virtually unlimited. As the double arrows for the working cylinder 10 and the hydraulic motor 12 symbolically show, both the working motion of the piston rod unit of the working cylinder 10 and also the respective direction of rotation for the hydraulic motor 12 can be reversed. Here it applies to the hydraulic motor 12 that, when the slewing gear 16 is being driven in one direction, upon acceleration it experiences a compressive load, while in braking a pulling load is formed because the inert mass (moment of inertia J) of the slewing gear 16 continues to move. The situation is comparable on the working cylinder 10 when a load is compressed in one direction and in the other, opposite direction must be pulled analogously for a retraction motion. Fundamentally, it then applies that when pulling loads move the respective consumer 10, 12 more quickly than corresponds to the volumetric flow in the, inlet channel flowing in on the utility ports A1, B1, A2, B2, the inlet pressure then can drop quickly to the cavitation pressure and below; this is to be avoided due to damaging effects. The controlling device which is to be detailed serves this purpose.
The controlling device for this purpose has a control valve 18 which among other purposes is used to control a supply line TReg for the respective hydraulic consumer 10, 12. Diametrically opposite to the output-side supply line TReg, the tank return line TR0 is also connected to the control valve 18 on the input side. Another output of the control valve 18 is connected to an additional feed line TR, and the control valve 18 is designed as a priority valve such that the supply line TReg acquires preference of fluid supply over the tank return line TR0.
A hydraulic pump means of conventional design, which is not detailed, is used for fluid supply or pressure supply p. The pressure supply p in turn is connected by way of a throttle D1 to the input side of the control valve 18 and the pressure supply p discharges into a secondary branch 20 to the input side of two other control valves 22, 23, one control valve 22 on the output side being connected with its fluid ports to the utility ports A1, B1 of the hydraulic cylinder 10 and the second, other control valve 24 is connected analogously to the utility ports A2, B2 of the hydraulic motor 12. The respective valve 22, 24 is moreover connected on the input side in a fluid-carrying connection to the feed line TR and the two outputs which lead to the respective utility ports A1, B1, A2, B2 are connected by way of a fluid line to the supply line TReg.
Two check valves 26, 28 at a time are connected into the pertinent fluid lines, and the check valve 28 which leads to the utility port B1, B2 is to be provided with a pressure limitation function. Furthermore, all the check valves 26, 28 open in the direction of their respectively assignable utility ports A1, B1, A2, B2. The other, second control valves 22, 24 are designed as 4/3 directional control valves and are shown in their middle unactuated position in which the respective input side is separated from the output side. The respective 4/3 directional control valve can be controlled hydraulically or electrohydraulically in the conventional manner by way of opposing control ports al, b1, and a2, b2. These 4/3 directional valves can also be optionally replaced by other valve constructions and in addition to the illustrated working cylinder 10 and the hydraulic motor 12, there can be other consumers of the same type or different type. The control block 14 also can be used solely for controlling a hydraulic consumer 10 or 12.
As furthermore follows from
The controlling device show in
The indicated sensor circuit relieves the tank return line TR0 as long as a supply demand is not indicated. Otherwise the tank return line TR0 is throttled to a mechanically predetermined level which is dictated essentially by the spring force of the control spring 34. As long as there is no quantitative outflow in the supply line TReg the fluid is routed unthrottled into the tank return line TR0. If, in contrast, a supply stream flows out, then the control valve 18 continues to control the mechanically set pressure in the supply line TReg by its throttling the outflow sectional view to the free tank return line TR0, in this way at the same time raising the pressure in TR above that in the supply line TReg, and the fluid medium now having to flow into the supply line TReg through the check valve RV. To implement the sensor, the pressure area active at the time on the valve piston 32 of the control valve 18, which is designed as a compensator, is used. According to the circuit diagram shown in
The control valve is shown in detail in the following figures using various working positions.
Furthermore, the individual annuli 46, 48, 50, 52, and 54 are separated essentially fluid-tight from one another by way of piston segments 56, 58, 60, and 62. These piston segments 56, 58, 60, and 62 are widened in diameter relative to the remaining diameter of the valve piston 32 with the formation of active annular piston surfaces. In the left-hand stop position shown in
Furthermore, between the first piston segment 56 and the second piston segment 58 there is a transverse hole 70 which discharges into the second annulus 48 in the illustrated working position as shown in
Furthermore, the transverse hole arrangements which follow one another in the longitudinal plane each can be located running offset to one another by 90° adjacently to one another. As furthermore follows from
If the compressive force acting on the control piston 32 becomes higher than the force of the control spring 34, the control piston 32 moves against the spring 34. In the process, the connection from TR to TR0 is opened and the feed pressure TR and the pressure in the supply line TReg drop. At this point, a control motion begins with the objective of setting the pressure in TReg exactly to the force of the control spring 34 as the valve spring. The pressure in the feed line TR can therefore not drop below the corresponding pressure value of the control spring 34 so that the return line TR0 is always preloaded.
In the already addressed theoretical operating sequence, the throttle D1 is now to be opened. The TReg fluid-carrying channel as the supply line is then supplied not only by way of the valve RV by TR, but also by way of the throttle D1, proceeding from a high pressure level, for example, in the form of the pump supply pressure p. If at this point no pulling loads occur, the valve piston 32 moves in turn, viewed in the direction of looking at the figure, to the right against the spring 34 and, in doing so, within the valve a fluid-carrying connection from TR to the unpressurized return line TR0 occurs. This causes the pressure in TR to drop. The valve RV closes because the supply line TReg is supplied additionally by way of the throttle D1 and a volumetric flow does not escape. Therefore the pressure in the supply line TReg remains at a level which corresponds to the amount of the control spring 34.
The control piston or valve piston 32 can then run completely against the control spring 34 without feed pressure being taken from TR. Therefore, when the connection TR to TR0 is completely opened, the pressure at TR drops to the level of TR0. If the valve piston 32 were to run against the spring 34 as far as the mechanical stop, the pressure in the supply line TReg would run to the level of the inlet pressure in the second feed line 30. This valve state is shown in
When pulling loads act on the respective consumer 10, 12, the control valve 18 as the tank back pressure valve blocks the outflow into the free tank return line TR0 and the control edge 76 of the second feed line 30 is completely opened. So that at this point some arbitrary amount of volumetric flow is not taken from the high pressure level, the inlet of the second feed line 30 is safeguarded with the throttle D2. To prevent the control motions from leading to vibrations, there is an opened third throttle site D2′ in the relief line 36. This results in a small control oil loss from TReg to TR which, however, as indicated, acts on the control valve 18 in a stabilizing manner and is negligibly small with respect to the implemented energy savings effects. The operating diagram as shown in
The level of control oil consumption is determined by the level of the force of the control spring 34 and the throttle action of D2. Typical design specifications vary between pressure preloading of 10 bar, combined with D2 equal to 0.8 mm and pressure preloading of 7 bar, obtained from two successively connected throttles of 0.6 mm as D2. Thus the control oil consumption of 1 l/min to 0.34 l/min can be easily varied for each design. The assigned energy losses are then dependent on the current pump pressure p which supplies the second feed line 30 at the same time. For an average pump pressure of 200 bar then losses of 0.3 KW and 0.1 KW occur.
If for some reason a volumetric flow is additionally supplied to the supply line TReg, for example, originating from the pressure limitation valves on the pipe ports, this additional volumetric flow must not lead to impermissible pressure piling in the supply line TReg. Therefore this additional volumetric flow must be reliably discharged. For this purpose the tank back pressure valve in the form of the control valve 18 can route its stroke until a connection of the supply line TReg to the feed line TR opens, the feed line TR being connected already unthrottled to the free tank return line TR0.
With the described controlling device especially the operating capacity of the hydraulic consumers 10, 12 can be increased if there is no supply state. A pressure increase by 7 to 10 bar is easily possible so that in this respect energy is also saved by this isolated back pressure. The assignable cooling system can also be made smaller due to this saving of energy. Furthermore, fuel is saved, particularly diesel fuel. The operating capacity of the hydraulic consumer, if there is no supply state, is increased, for example, by 7 to 10 bar. Then the energy is saved in the amount of the isolated back pressure of 7 to 10 bar. One typical example is a small excavator with an average volumetric flow of 50 ml/min and 7 bar back pressure in the supply channel. This yields energy savings of approximately 0.58 kW, of which, however, approximately 0.1 KW control oil loss would have to be subtracted. For a mobile excavator with an average volumetric flow of 200 l/min at a back pressure of 10 bar, even 3.3 KW could be saved. Of this, a maximum of 0.3 KW would again be lost for control oil loss so that the total energy savings would be 3 KW.
| Number | Date | Country | Kind |
|---|---|---|---|
| 10 2006 061 305.8 | Dec 2006 | DE | national |
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/EP2007/009295 | 10/26/2007 | WO | 00 | 11/20/2009 |
