Patent Application
20130042935
The invention relates to a hydraulic switching mechanism for the mobile hydraulics of mobile hydraulic machines, in particular hydraulic excavators, with a valve block, with electrohydraulically activatable valve units arranged in the valve block for controlling the movement of a working cylinder having two oppositely acting cylinder chambers which can in each case be connected via cylinder connections to the valve block, wherein the cylinder connections can be selectively connected to a pump connection for hydraulic fluid, to a tank connection or to one another, and with pre-control valves for the electrohydraulic activation of the valve units, wherein a directional control valve function for the direction of the movement, and a lowering braking function for the sequence of the movement, of the associated working cylinder can be controlled by means of the hydraulic switching mechanism. The invention also relates to mobile hydraulic machines having such a hydraulic switching mechanism and to valve units therefor.
In the case of driveable and hence mobile working machines, the particular constraints and demands placed on the structural design of the hydraulic devices have resulted in the independent category of mobile hydraulics being developed in parallel to stationary hydraulics, and the invention relates to the technical field of mobile hydraulics. In hydraulic drives for controlling a hydraulic cylinder or hydraulic motor, the drive movement normally occurs with pressure and throughflow generated in a pump unit against the load forces acting on the cylinder from outside counter to the direction of movement (positive load forces). However, it is also possible in the course of movement for negative load forces to occur in the direction of movement—such as during lowering of lifted loads, a braking of moved masses or load direction reversal—which result in undesired leading and uncontrolled lowering of the cylinder. In addition to the uncontrolled movement, a negative pressure with cavitation would occur on the cylinder side driven by the pump throughflow, with the result that the hydraulic system may be damaged. In order to control the working cylinders in mobile hydraulic machines, use is made of 6/3-port directional control valves of piston slide valve type with a proportional throttling function which are designed specifically for use in mobile hydraulics and which, upon activation, throttle in a proportionally controlled manner both the oil inflow from the pump to the working cylinder and the oil outflow from the working cylinder to the tank. The main working movements—generally during the extension of the cylinder—occur with positive force loadings, wherein the load acts in a pushing manner counter to the desired direction of movement of the consumer. However, negative force loadings can also occur in both directions of movement, wherein the load acts in a pulling manner in the same direction as the desired direction of movement, such as, for example, during the lowering of loads, braking of large moved masses and load change of externally acting forces. As a consequence, the volumetric flow flowing from the cylinder to the tank must be throttled in order to prevent undesired acceleration and uncontrolled movement of the cylinder, and it is known to provide valves having a lowering braking function for this purpose. In mobile hydraulics, use is made of complex mobile control blocks having a plurality of 6/3-port directional control valves with all the required additional functions, including the throughflow distribution to the connected cylinders from a delivery pump.
Excavator booms and other working manipulators, such as shovels, buckets or sliding ploughs, within the sector of mobile working machines are nowadays predominantly controlled by the operator by means of hand lever pre-control devices (joysticks). When problematic operating states occur, which may be caused for example by changing loads or particularly quick or slow movements, the operator must then perform a corresponding actuating signal correction to maintain the desired setpoint movements, something which requires appropriate training and experience. With regard to boom and dipper cylinder control of a shovel excavator, there is obtained a separate function whereby, after the extension operation, the lowering during the retraction of the cylinders is intended to take place through self-weight without pump inflow. This function is referred to below as “floating”. For this purpose, the piston side and rod side of the working cylinder are bypass-connected or short-circuited. The oil displaced from the piston side through the force of the weight flows, in order to replenish the oil volume sucked away, partially to the rod side and the residual quantity flows to the tank. The lowering speed is electrohydraulically proportionally controlled by a throttling bypass valve in a variable remote-controllable manner. The residual quantity flowing to the tank flows via a pre-stressing return valve which pre-stresses the pressure in the cylinder connection to such an extent that no cavitation can occur in the cylinder through flow losses in the cylinder line. These valves which are required for the lowering in bypass mode through self-weight must additionally be installed in the main flow with corresponding throughflow capacity between the mobile control block and cylinder. Since the mobile hydraulics used to date produce a throughflow in the part-load range via a bypass, there occur considerable hydraulic energy losses which considerably reduce the efficiency of the drive and require a large cooling capacity of the hydraulic system. This loss effect occurs particularly when braking negative load forces in the direction of movement since, in order to throttle the throughflow flowing back from the cylinder, the hitherto used valve units with valve slides have to be actuated in the closing direction always in the fine-control range with control edge undercutting. These hydraulic energy losses caused by the valve control principle come increasingly to the fore as a disadvantage as the overall size and drive power of the mobile working machine increases.
In particular in the case of large mobile machines and large-area excavators as are used, for example, in open-cast mining, given the high loads to be controlled, the required throughflow quantities and throughflow rates of far above 1000 L/min (264 gal/min), and the aforementioned disadvantages, mobile machines with a cable control are usually used.
An object of the invention is to provide a hydraulic switching mechanism for mobile hydraulics that does not have the aforementioned disadvantages, can be operated with fewer hydraulic energy losses and makes it possible to dispense with cable controls even in the case of large hydraulic machines.
These and further objects are achieved by a hydraulic switching mechanism for the mobile hydraulics of mobile hydraulic machines, in particular hydraulic excavators, with a valve block, with electrohydraulically activatable valve units arranged in the valve block for controlling the movement of a working cylinder having two oppositely acting cylinder chambers which are connectable via a cylinder connection to the valve block, wherein the cylinder connections are selectively connectable to a pump connection for hydraulic fluid, to a tank connection or to one another, and with pre-control valves for the electrohydraulic activation of the valve units, wherein a directional control valve function for the direction of the movement, and a lowering braking function for the sequence of the movement, of the associated working cylinder is controllable by way of the hydraulic switching mechanism. The mechanism further including four cone-seat valve units each having a spring-loaded valve cone in the valve block for the working cylinder, of which the first forms a pump valve unit between the first cylinder chamber connection and the pump connection, the second forms a tank valve unit between the first cylinder chamber connection and the tank connection, the third forms a pump valve unit between the second cylinder chamber connection and the pump connection, and the fourth forms a tank valve unit between the second cylinder chamber connection and the tank connection, wherein a pressure-limiting function and the lowering braking function are realizable for both directions of movement in a pressure-dependent manner as a function of the pressure in the cylinder chamber connections by way of the tank valve units via an associated pre-control valve system including a plurality of pre-control valves. Further advantageous configurations, specific solutions for the main application area of large hydraulic machines, and also valve units which can be used with advantage, are indicated in the remaining disclosure of this application.
Provision is made according to the invention for four cone-seat valve units comprising cone seat valves and each having a spring-loaded valve cone to be provided in the valve block for a working cylinder, of which the first valve unit forms a pump valve unit between the first cylinder chamber connection and the pump connection, the second valve unit forms a tank valve unit between the first cylinder chamber connection and the tank connection, the third valve unit forms a pump valve unit between the second cylinder chamber connection and the pump connection, and the fourth valve unit forms a tank valve unit between the second cylinder chamber connection and the tank connection, wherein a pressure-limiting function and the lowering braking function can be achieved for both directions of movement in a pressure-dependent manner as a function of the pressure in the cylinder chamber connections by means of the tank valve units via an associated pre-control valve system comprising a plurality of pre-control valves. In the case of the hydraulic switching mechanism according to the invention, the control block is provided with four valve units having cone seat valves optionally designed for maximum throughflow rates in order to control the working cylinders with the directional control valve functions of starting, stopping and direction of movement control and, by means of a suitable pre-control valve system, also lowering through weight loading in cylinder bypass control without additional valves, it being possible, as a function of the pre-control valve system, for the tank valve units to be given additional valve functions such as directly controlled with superimposed pre-controlled lowering braking function, maximum pressure safeguarding of the cylinders, and proportional throttle valve function for the controlled displacement under negative load forces in the direction of movement and braking during an emergency stop. Particularly for hydraulic excavators for moving large loads, it is advantageous if, in order to achieve optimum energy utilization, the speed control of the working cylinder movement occurs directly by adjusting the pump delivery flow without additional throttle valve functions. The cylinder connections can each be connected to a pump unit via the two pump valve units. The cylinder connections can each be connected to the tank via the tank valve unit. The valve cones of the tank valve units are controlled and positioned pressure-dependently via a control connection and also via the pilot and pre-control valves which are preferably integrated in a valve block.
To optimize the mobile hydraulics, it is particularly advantageous if the tank valve units make it possible, in addition to the directional control valve function for starting, stopping and direction influencing, to ensure a blocking function in the zero position, maximum pressure safeguarding of the two cylinder chambers, hence a piston side or a cylinder rod side of the working cylinder, a counterpressure function with adaptation of the counterpressure to the cylinder load force, hence a lowering braking valve function with activatable, relievable counterpressure function for both directions of movement of the working cylinder, and an electrohydraulic proportional throttle valve function for the cylinder outflow control to the tank during the braking of negative cylinder load forces and moved masses independently of the delivery flow control of the pumps. It is further advantageous if the proportional throttle valve function can also be used in addition to controlling the lowering operation for the cylinder retraction through cylinder load force (weight force) without pump inflow, i.e. a so-called “floating”, something which can be achieved, in particular, if, according to a particularly advantageous configuration according to the invention, the proportional throttle valve function is integrated via the pre-control valve system into both tank valve units. The combining of a plurality of valve functions in a valve unit correspondingly requires a pre-control circuit, composed of a plurality of pilot or pre-control valves, in a pre-control valve system, and the text which follows reveals numerous advantageous configurations and variants of valve units and pre-control valve systems for achieving the plurality of valve functions in combination with a compact and operationally reliable construction of the hydraulic switching mechanism.
According to an advantageous configuration, the valve cones of the tank valve units can have a seat surface which is directly pressurized with the pressure in the associated cylinder connection, and a control surface which is indirectly pressurized with the same pressure through the interposition of a pressure-limiting valve in the pre-control valve system. The switching position of the valve cone is dependent on the control pressure exerted on the control surface in relation to the pressure forces which are active on the seat surfaces via the hydraulic pressure in the cylinder chamber connections. When the control pressure is relieved, the valve cone opens and throughflow can occur in both directions; when the control pressure is applied, the valve cone closes and blocks the throughflow in a leakage oil-free manner. Further preferably, a nozzle can be arranged in a control line between the cylinder connection and the pressure-limiting valve, and/or a nozzle can be arranged in a control line between the pressure-limiting valve and a control chamber for pressurizing the control surface. The tank valve units can then form pressure-limiting valves which are pre-controlled in their output function, it being possible by switching a pre-control valve in the pilot valve system to achieve additional pressure relief.
In order to increase the opening pressure of the valve cone to blocking pressures of, for example, 60 (870 psi) bar to 100 bar (1450 psi), as may occur in particular when using the hydraulic switching mechanisms according to the invention in the mobile hydraulics of heavy-load excavators, in addition to a valve spring, the valve cone of the tank valve unit can be subjected to the spring force of a disc spring stack in the direction of the valve seat. According to a particularly advantageous configuration, the valve cone is designed as a hollow socket with a cavity situated opposite the seat surface, wherein the valve spring and a plunger each bear against the valve cone by one end at the bottom of the cavity, and the other end of the plunger is subjected to the spring force of the disc spring stack. The installation into the tank valve units of a disc spring column guided by a plunger means that it is possible, via the plunger, to transmit additional high closing forces to the valve cone, and an additional directly controlled pressure limiting can be produced. Consequently, and as a result of the friction between the series-arranged disc springs, there is an improvement in the stability in the regulating response in the case of the installed pre-controlled pressure functions. The directly acting closing function of the disc spring stack on the valve cone affords an additional safety function, which means that even in the event of a failure of the pre-control system—for example in the event of clogging of the inlet nozzle to the pressure-limiting valve and resulting lack of pressure build-up on the valve cone control surface, this directly acting counterpressure of the disc spring force remains for braking purposes.
It is particularly advantageous if, in the case of the tank valve units, a lifting piston sleeve with a lifting piston is arranged between the disc spring stack and the valve cone, wherein that surface of the lifting piston which is situated facing away from the disc spring stack forms a lifting piston control side and can be subjected to or is subjected to the hydraulic pressure of the respective other cylinder chamber connection via a control line. Preferably, the lifting piston is guided displaceably on the plunger and is moveable relative to the plunger in the axial direction. This lifting piston function is mechanically kinematically uncoupled from the valve cone/plunger movement and acts only on the column of the disc spring stack, with the result that closing and pressing functions with the valve cone which are controlled by the valve pre-control system are possible in parallel and at the same time. There results the function of a directly controlled lowering braking valve with an activatable counterpressure function.
In order to achieve an extended lowering braking function even for higher load-holding pressures of up to about 350 bar (5,076 psi), according to an alternative embodiment, there can be arranged in the valve pre-control system a directly controlled pilot lowering braking valve with a valve cone slide which has an opening pressure surface which is subjected, via a preferably electrically activatable pilot valve, to the pressure of the control line connected to the associated cylinder chamber connection, and which has a pressure activation surface which is subjected, via a pressure return line, to the pressure in the other cylinder chamber in order to bring about an additional pressure relief at the control pressure surface of the valve cone. Upon actuation of a pilot valve, this directly controlled lowering braking valve with an activatable counterpressure function as a pre-control valve for controlling the pressure of the valve cone of the tank valve unit is switched on. The tank valve unit then operates in the basic function as a hydraulically pre-controlled lowering braking valve. The pre-control valve can be set to the maximum load-holding pressure of the respective application with an additional safety of 20-30% so that this cylinder load is securely blocked against undesired lowering. Through the pressure return, the pilot lowering braking valve opens at substantially lower pressures than the set maximum load-holding pressure and, at the pressure control surface of the valve cone, generates a lower control pressure which, together with the directly controlled lowering braking function with valve cone and disc spring stack, produces a resulting braking counterpressure on, for example, the cylinder piston side. Even under a changing negative load force, this braking counterpressure still remains precisely high enough for a low drive pressure to be permanently established on the rod side of the working cylinder. The directly controlled lowering braking function with valve cone, plunger, lifting piston and disc spring stack is always active when there is a low drive pressure on the rod side (or the piston side). The pre-controlled lowering braking function is only activated when there is a high drive pressure on the rod side in order, under a simultaneously negative load force, to produce the required counterpressure for a controlled lowering via the control pressure on the valve cone. With a load change and a positive load force F against the direction of movement, it is possible as a result of the required high driving pump pressure on the cylinder rod side for the disc spring stack to be raised by the lifting piston as far as a lifting piston stop such that this stack no longer acts on the valve cone. At the same time, the pilot lowering braking valve can be completely activated and the control pressure on the valve cone can be completely removed to the tank such that the valve cone opens against the valve spring as a non-return valve, with the result that a counterpressure braking the retraction movement is avoided on the piston side. In the event of a sudden stop in an emergency situation, it is also possible, independently of the lifting piston/disc spring stack assembly, for the valve cone to be displaced into the closed position by relieving the pilot directional control valve. The influence of the pre-controlled lowering braking valve function can be varied through the use of interchangeable pilot lowering braking valves with different transmission ratios by means of stepped pressure activation surfaces and thus adapted to the different conditions of the overall control. Further adaptation of the effect of this pre-controlled lowering braking valve function is possible via the size of a nozzle preferably connected upstream of the pilot lowering braking valves. The directly controlled lowering braking function with valve cone and disc spring stack and lifting piston for counterpressure control leads to a significantly improved stability behaviour.
In order to be able in a simple manner to modify the transmission ratio for the direct lowering braking function in order to reduce the drive pressure, it is advantageous if the lifting piston is installed in an interchangeable insert which can be interchanged as a structural unit in a completely functional manner after disassembly of a valve block cover and, if appropriate, can be replaced by lifting pistons having different hydraulic active surfaces.
According to a further advantageous configuration, a proportional throttle valve function is possible with the hydraulic switching mechanism. The additional proportional throttling function can be controlled in particular via the tank valve units and the pre-control system for regulating the hydraulic oil flow from the cylinder to the tank connection. The proportional throttling function ensures that a “floating”, i.e. a control of the lowering movement through self-weight without pump inflow for the cylinder retraction, is possible, a limiting of the maximum cylinder speed is ensured in the case of delayed response of the lowering braking valve function and/or in the case of extreme cylinder load conditions, and furthermore a proportionally controlled outflow throttling function is made possible during load cycles with stability problems occurring during the lowering braking function. In a normal case, the lowering movement of the cylinders should here take place through the weight force acting on the cylinder as a negative load force in the direction of movement. By activating further pilot directional control valves, the two pump valve units C1 and C3 can be opened and the cylinder chambers of the working cylinder, hence the piston side and rod side of the working cylinder, can be hydraulically connected. If at the same time a tank valve unit is opened in a throttled manner, a portion of the throughflow displaced from the piston surface flows, corresponding to the surface ratio of the cylinder, via the pump valve unit arranged in series for this purpose in order to replenish the oil volume sucked away from the cylinder rod side ZB. The remaining residual flow displaced as surplus flows away in a throttled manner to the tank, with the lowering speed of the cylinder being determined by setting the throttling opening cross section. A return flow to the pump is preferably prevented by a non-return valve in the pump inlet. Since the weight force acts directly on the piston rod surface after the short-circuit connection of the cylinder connections, owing to the resulting higher pressure through pressurization of the lifting piston via the control line Z2, this lifting piston will raise the disc spring stack and completely cancel or at least to a large degree compensate for the closing force on the valve cone.
The opening stroke of the valve cone of the tank valve units, which stroke is proportional to a predetermined electrical signal, can be produced by means of different electrohydraulic positioning systems. For the preferred application area of hydraulic excavators which are exposed to harsh environmental influences, simple, robust systems without electronics installed on the valve are preferred for internal return lines. According to an advantageous variant embodiment, it is possible, in particular to set the throttling opening cross section, for the tank valve units each to be assigned an adjusting piston system with internal position regulation through force balancing. The actuating piston system is preferably arranged in a portion adjoining the installation chamber for the disc spring stack and comprises a pressurized actuating piston which bears against the plunger with pre-stressing in the closing direction of the valve cone. The actuating piston preferably has a pressure surface which is larger, preferably about 1.1 to 2.2 times larger, than the seat surface of the valve piston of the assigned tank valve unit. The pressurization of the actuating piston is preferably adjustable by means of a proportional magnet, a control piston and a return spring and/or the actuating piston system is assigned a control valve with alternate pump connection or tank connection coupling. The proportional throttling function can then become operative in a superimposed manner with respect to the opening stroke limiting during the lowering braking function and separately also as outflow throttling during cylinder lowering through self-weight (floating), wherein the closing force of the disc spring stack is reduced or cancelled corresponding to the cylinder pressure which is established after the connection of the two cylinder sides. By virtue of the force-locking connection of the valve cone via the plunger against the actuating piston with an enlarged active pressure surface, there results a differential piston assembly which can be positioned by pressure control on the actuating piston surface acting in the closing direction via a three-way control valve having alternate pressure or tank connection. The positioning is carried out in the closed position control loop by force balancing at the control piston between the actuating force of the proportional magnet as a set point value and the spring force, produced by the actuating piston proportionally to the opening stroke, of a return spring as actual value. Alternatively, in order to control the throttling opening cross section, the tank valve units can each be assigned an electric stepping motor, in particular a linear motor, and a following piston system comprising a control piston and following piston. The positioning of the opening stroke by the proportional throttle valve can continuously occur analogously through adjustment of the control valve by the proportional magnet or the electrical linear motor during the lowering movement. However, the stroke opening position can also be set as a fixed set point value at the proportional magnet or the electrical linear motor before the lowering movement. Upon actuation of an assigned pilot directional control valve, the adjusting piston or following piston, coupled with the valve cone, runs into this predetermined position.
The pilot control valves and pilot directional control valves of the pre-control valve system and/or the overall pilot control circuit are preferably arranged in a valve housing cover which can be releasably connected to the valve block.
The main application area of the invention concerns hydraulic machines, in particular large hydraulic excavators having flow rates far in excess of 1000 L/min (264 gal/min), with at least one hydraulic cylinder as working cylinder for adjusting at least one arm connected to a working implement such as a bucket, shovel or the like, with a pump unit for generating a hydraulic oil flow, with a hydraulic switching mechanism comprising a valve block as mobile hydraulics for the hydraulic machine, with electrohydraulically activatable valve units arranged in the valve block for controlling the movement of the working cylinders, and with pre-control valves in the hydraulic switching mechanism for the electrohydraulic activation of the valve units, wherein a directional control valve function for the direction of the movement, and a lowering braking function for the sequence of movement, of the associated working cylinder can be controlled by means of the hydraulic switching mechanism, wherein a hydraulic switching mechanism designed according to the invention, as described above, is used in these hydraulic machines. It is then particularly advantageous if, in particular to achieve optimum energy utilization in all load ranges with particular consideration to part load, the speed control of the working cylinder is performed only via the pump delivery flow without additional control valve throttling losses. For this purpose, when use is made of diesel engines as the drive unit, the pump delivery quantity can be produced with variable displacement pumps and, by electrohydraulic adjustment of the pivoting angle, the delivery flow and hence the speed of the working cylinders can be controlled. Additional throttle valves in the cylinder inflow with energy losses for controlling the delivery flow regulator of the variable displacement pump are then no longer required. When use is made of electrical three-phase motors as the drive unit, the pump delivery flow can be produced with fixed displacement pumps and be regulated by rotational speed regulation with frequency converters.
The invention also relates to the valve unit for the above-described hydraulic switching mechanism for mobile hydraulic machines, in particular tank valve units which are designed as a cone-seat valve of cartridge construction which can be inserted into a bore in the valve block and which comprises a valve sleeve, valve cone and valve spring, wherein the valve cone is designed as a hollow socket with a cavity situated opposite to a seat surface as a bearing surface for the valve spring and for a plunger which is subjected to or can be subjected to the spring force of a disc spring stack. It is particularly advantageous if the disc spring stack and the plunger are arranged together with a lifting piston in a lifting piston sleeve, wherein the lifting piston is guided displaceably on the plunger and is moveable relative to the plunger in the axial direction of the mounting bore in the valve block, and that side of the lifting piston which is situated facing away from the disc spring stack forms a lifting piston control side. The lifting piston sleeve together with the associated functional parts can be advantageously arranged in the valve block cover so that, by exchanging the lifting piston sleeve for a lifting piston sleeve having different active surfaces and/or by exchanging the cartridge valves for a cartridge valve having a different valve nominal size, optimum adaptation to the required throughflow capacities and pressure conditions can be achieved.
Further advantages and configurations of a hydraulic switching mechanism according to the invention, particularly for use in large hydraulic machines, will become apparent from the description given below of schematic figures for the construction of the switching mechanism together with the associated pilot valve control circuit.
Further, these and other objects, aspects, features, developments and advantages of the invention of this application will become apparent to those skilled in the art upon a reading of the Detailed Description of Embodiments set forth below taken together with the drawings which will be described in the next section.
The invention may take physical form in certain parts and arrangement of parts, a preferred embodiment of which will be described in detail and illustrated in the accompanying drawings which form a part hereof and wherein:
Referring now to the drawings wherein the showings are for the purpose of illustrating preferred and alternative embodiments of the invention only and not for the purpose of limiting same,
To achieve optimum energy utilization in all load ranges with particular consideration to the part load, the speed of the operating cylinders 6 in a hydraulic excavator 1 designed according to the invention is controlled only via the pump delivery flow of the pump 16 without additional control valve throttling losses. When a diesel engine is used as the drive unit for the hydraulic excavator 1, the pump delivery quantity is generated with variable displacement pumps, with the delivery flow and hence the speed of the working cylinders being controlled by electrohydraulic adjustment of the pivoting angle. Additional throttle valves in the cylinder inflow with energy losses for controlling the delivery flow regulator of the variable displacement pump are then no longer necessary. When electrical three-phase motors are used as the drive unit, the pump delivery flow can be generated with fixed displacement pumps and be regulated by rotational speed regulation with frequency converters.
Via the two pump valve units C1, C3, the cylinder chambers ZA and ZB can be respectively connected to the pump unit 16 or to the pump connection P via the associated cylinder connections A, B in the valve block 5. Via the two tank valve units C2, C4, the cylinder chambers ZA and ZB can be respectively connected to the tank via the tank connection T. As has been shown specifically for the tank valve units C2, C4 in
The cylinder 6 is extended during operation with a signal preset at the hand lever (joystick) 15 by proportional delivery flow setting at the pump unit 16 for setting the speed and simultaneous actuation of the directional control valve function by opening of pump valve units C1 and tank valve units C4 during activation by the electrical pilot controller 17 of the pilot directional control valves arranged in the valve block cover or covers 26, with the result that the control surfaces 31 in the control oil chamber 32 are pressurelessly relieved and the valve cones 28 open while being pressurized by the main flow connections. The working cylinder 6 is retracted with pump inflow by activating and opening pump valve unit C3 and tank valve unit C2.
For the floating function for lowering the working cylinder 6 through self-weight without pump inflow, the two pump valve units C1, C3 are opened for bypass-connection of the cylinder connections ZA with ZB. By opening the tank valve unit C4 equipped with an additional proportional throttling function for controlling the lowering speed, the excessively displaced residual oil quantity flows to the tank.
All of the directional control valve functions required for the cylinder control are carried out by the four cone-seat valve units C1, C2, C3 and C4 arranged in the mobile valve block 5. Each of these cartridge valves can be optimally adapted to the required throughflow arrangements by selecting the valve nominal size, for which reason a parallel connection of valves to achieve the throughflow capacity, as previously employed in the prior art, is dispensed with.
The valve block covers 26 contain all the pilot valves required to control the respective valve unit C1, C2, C3 and C4 in order to relieve the mobile valve block 5 of control bores.
In addition to the directional control valve function for starting, stopping and direction influencing, the tank valve units C2, C4 contain the following valve functions via the construction of the pilot control system 7:
The construction of the tank valve units and of the pilot valve system for implementing the aforementioned valve function will now be explained with reference to the further figures. The combining of a plurality of valve functions is achieved by means of a pre-control circuit 7 for the tank valve units C2, C4 which is composed of a plurality of pilot valves and which is integrated substantially completely into the valve block cover 26. The fundamental overall construction of the tank valve units C2, C4 can be seen from
In the basic position according to
In order with the tank valve units C2 and C4 to ensure a maximum pressure safeguarding, a lowering braking valve function and a superimposed electrohydraulically actuated throttle valve function, the tank valve units, as shown in
A description will now be given first, with additional reference to
This directly controlled lowering braking function with the valve cone and disc spring stack can, given the overall size of the spring, only be meaningfully carried out up to maximum blocking pressures of about 60-100 bar (870-1450 psi). Therefore, this directly controlled lowering braking function is extended and supplemented for higher load-holding pressures up to 350 (5076 psi) bar in that an additional lowering braking function with a pilot lowering braking valve of smaller overall size is integrated into the hydraulic pre-control circuit 7. The simplified scheme of this cartridge embodiment composed of directly and additionally pre-controlled lowering braking function is represented in
The driving pump pressure in the cylinder chamber ZB on the cylinder rod side that is required for retracting the cylinder despite negative force action in the retraction direction is applied, through the pressure return via the connection Z2, nozzle NZ2 and activation connection 23 which are arranged or formed in the valve block cover 26, to the additional pressure activation surface 21 of the pilot lowering braking valve PCB. This valve opens at considerably lower pressures than the set maximum load-holding pressure and generates, in the control oil chamber 32 of the valve cone 28, a lower control pressure which, together with the directly controlled lowering braking function with the valve cone 28 and disc spring stack 36, brings about a resulting braking counterpressure pZA in the cylinder chamber ZA on the cylinder piston side. Even with a changing negative load force F, this braking counterpressure pZA is precisely still high enough for a low drive pressure pZB to be established in the cylinder chamber ZB on the rod side of the cylinder 6. The directly controlled lowering braking function with valve cone 28, plunger 34, lifting piston 38 and disc spring stack 36 is always active when there is a low drive pressure pZB in the cylinder chamber ZB on the rod side. The pre-controlled lowering braking function is only activated when there is a high drive pressure pZB in the cylinder chamber ZB on the rod side in order, with a simultaneously negative load force, to generate the required counterpressure in the cylinder chamber ZA for controlled lowering via the control pressure for the valve cone 28. The lifting piston 28 has then completely relieved the disc spring stack 36, as shown in
With a load change and a positive load force F against the direction of movement, as shown in
The lifting piston function also serves for compensating for or cancelling the disc spring closing force for the electrohydraulic proportional throttle valve function. The additional proportional throttling function at the tank valve units C2, C4 from the cylinder 6 to the tank return T allows a control of the lowering movement through self-weight without pump inflow for the cylinder retraction (floating), a limiting of the maximum cylinder speed with a delayed response of the lowering braking valve function and/or in extreme cylinder load conditions and forms the precondition for a proportionally controlled outflow throttling function during load cycles with stability problems occurring during the lowering braking function.
The lowering movement of the cylinders should in the normal case occur through the weight force acting on the cylinder as a negative load force in the direction of movement. By activating the pilot directional control valves PVC1 and PVC3, the two pump valve units C1 and C3 are opened, as shown in
The opening stroke of the valve cone 28 which is proportional to a predetermined electrical signal can be produced using various electrohydraulic positioning systems. For use in mobile hydraulic excavators which have to operate under harsh environmental influences, simple robust systems without electronics installed on the valve are preferred for internal return lines, and two advantageous positioning systems will now be described with reference to
In the valve block cover 26 according to
Alternatively, the electric actuating signal for the proportional throttling function can be converted into a linear actuating travel by an electric stepping motor or servo motor via a threaded spindle, and a mechanical-hydraulic following piston system can hereby be activated for force amplification. The construction of this following piston system can be seen from
The foregoing description will reveal to a person skilled in the art numerous modifications that are intended to come within the scope of protection of the appended claims. The figures merely show advantageous exemplary embodiments without limiting the scope of protection of the appended claims. In the case of hydraulic excavators and other hydraulic working machines, a plurality of working cylinders must usually be operated partly simultaneously and partly successively, which is why a hydraulic switching mechanism usually comprises a plurality of valve blocks having the above construction.
Further, while considerable emphasis has been placed on the preferred embodiments of the invention illustrated and described herein, it will be appreciated that other embodiments, and equivalences thereof, can be made and that many changes can be made in the preferred embodiments without departing from the principles of the invention. Furthermore, the embodiments described above can be combined to form yet other embodiments of the invention of this application. Accordingly, it is to be distinctly understood that the foregoing descriptive matter is to be interpreted merely as illustrative of the invention and not as a limitation.
| Number | Date | Country | Kind |
|---|---|---|---|
| 10 2009 025 827.2 | May 2009 | DE | national |
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/IB10/52094 | 5/11/2010 | WO | 00 | 6/8/2012 |
