Abstract:
A control circuit, in particular for a slewing gear of a digger, has a hydraulic fluid tank ( 21 ), connected to two adjusting pressure chambers ( 10, 11 ). Each of the connections contains a separate brake valve ( 19, 20 ). The first of the valves ( 19 ) is operated by the differential between the working pressure in the first working conduit ( 2 ), and the control pressure in the control conduit ( 33, 34 ) charged with the higher pressure. The second valve ( 20 ) is operated by the differential between the working pressure in the second working circuit ( 3 ) and the control pressure in the control circuit charged with the higher pressure. Slow braking by the brake valves is interrupted when the slewing gear swings out against a resistance such as a heap of debris.

Description:
FIELD OF THE INVENTION 
     The invention relates to a hydraulic control system, in particular for activating a slewing gear of a digger. 
     A hydraulic control system is known from DE 196 20 664 C1. In the slewing gear control system disclosed by said publication, an adjusting apparatus is provided for adjusting an actuating piston, which is disposed between two actuating pressure chambers and influences the displacement volume of a hydraulic pump. The adjustment of the actuating piston is effected in dependence upon the pressure difference between two actuating pressure lines, which are connected each to one of the actuating pressure chambers. The actuating pressure in the actuating pressure lines is predetermined by two control lines connected to a manual control transmitter. Provided in each actuating pressure line is a separate braking valve, which throttles the return flow of hydraulic fluid from the actuating pressure chamber associated with the braking valve into a hydraulic fluid tank and hence enables the slewing gear to swing slowly outwards after the manual control transmitter has been returned into its neutral position by the operator. The effect achieved by the use of two separate braking valves, which are each connected to one of the working lines which connect a hydraulic motor driving the slewing gear to the hydraulic pump so as to form a working circuit, is that the slow braking by the braking valves is interrupted when the stewing gear swings out against a resistance, e.g. a heap of debris. 
     A drawback of the known hydraulic control system is however that the braking valves are disposed in the actuating pressure lines and are therefore biased by the actuating pressure. During the excursion of the actuating piston for accelerating the slewing gear, the hydraulic fluid filling the appropriate actuating chamber therefore flows through the braking valves, which are therefore exposed to increased fouling. The discharge of hydraulic fluid to the hydraulic fluid tank is effected by means of the manual control transmitter over relatively long line paths. Thus, the return flow of hydraulic fluid is restricted not only by the throttle provided in the braking valve but also by the area of the control lines and the opening area of the manual control transmitter. As a result, the time constant for the return flow of hydraulic fluid from the actuating pressure chambers of the adjusting apparatus may be adjusted reproducibly only to a limited extent by the throttle area of the braking valves. In said case, it has to be taken into account that the line length of the control lines, the manual control transmitter used and further structural parameters vary depending on the type of digger in which the hydraulic slewing gear control system is to be installed. The throttle area of the braking valves therefore has to be individually adapted to each type of digger, which entails a high assembly outlay. In addition, because the throttle areas of the braking valves used in DE 196 20 664 C1 are not adjustable, an adjustment after installation is not easily possible. 
     A further slewing gear control system is disclosed by DE 196 20 665 C1. In said slewing gear control system, the actuating pressure for the actuating pressure chambers of the adjusting apparatus is derived from the supply pressure of a supply apparatus via one or two pressure control valves. In said case, only one common braking valve for both actuating pressure chambers is provided, which is disposed in return flow direction downstream of a pilot device or pilot valve. In said refinement also, the return flow of hydraulic fluid first passes through the pilot valve, which likewise throttles the return flow, before reaching the braking valve. The effective throttle area therefore depends not only upon the throttle area of the braking valve but also upon the throttle area of the pilot valve as well as the areas of the connecting lines. The adjustment of the effective throttle area for the return flow of hydraulic fluid and hence the adjustment of the braking of the slewing gear is therefore made more difficult with said construction of the slewing gear control system also, especially as a variable, adjustable throttle area for the braking valve is not provided. 
     The object of the invention is therefore to indicate a hydraulic control system, in particular for activating the slewing gear of a digger, whereby the throttle area for the return flow of hydraulic fluid through the braking valves is definable more precisely and fouling of the braking valves is moreover counteracted. 
     SUMMARY OF THE INVENTION 
     The invention is based on the discovery that it is advantageous to dispose the braking valves, without interposing further valves, directly between the actuating pressure chambers of the adjusting apparatus and the hydraulic fluid tank. The resultant effect is short line paths for the return flow of hydraulic fluid from the actuating pressure chambers to the hydraulic fluid tank via the braking valve so that the effective throttle area depends substantially upon the throttle area defined by the braking valve and only to an insignificant extent upon the line areas. In the return flow path, apart from the braking valve, no further valves effecting an additional throttling are provided. By virtue of the fact that only the return flow of hydraulic fluid passes through the braking valves but not the flow of hydraulic fluid into the actuating pressure chambers in the event of acceleration of the slewing gear, the fouling of the braking valves is markedly reduced. In order, in the event of swinging-out of the hydraulic pump and loading of the actuating pressure lines with actuating pressure, to prevent a hydraulic short circuit of the actuating pressure lines via the braking valves towards the hydraulic fluid tank and, on the other hand, prevent a reflux of the returning hydraulic fluid into the actuating pressure lines or control lines, in each case a control valve is disposed in return flow direction downstream of a branch leading to the respective braking valve. According to the invention, the control valves and the braking valves are activated in such a way by the control pressure prevailing in the control lines that in the event of swinging-out of the hydraulic pump the control valves open and the braking valves close and, conversely, the control valves close and the braking valves open into their throttled valve position when the hydraulic fluid flows from the actuating pressure chambers back to the hydraulic fluid tank. 
     It is advantageous to provide the throttle area of the braking valves in an adjustable manner. This becomes possible only by virtue of the solution according to the invention, namely the arrangement of the braking valves not in the actuating pressure lines but in secondary lines, which branch off to the pressure medium tank, are biased with a lower pressure and exposed to less fouling. The braking valves in the known hydraulic control system take the form of seat valves so that they can withstand the actuating pressure there and be less susceptible to fouling. With seat valves, the construction of an adjustable throttle area is impossible or possible only with difficulty. An adjustable throttle area can be constructed more easily with a slide valve. A slide valve cannot however be used in the known hydraulic control system because, in the event of fouling, it can jam and hence lead to serious malfunctions. Given the development according to the invention, the use of a slide valve in the secondary line leading to the hydraulic fluid tank is however possible. In said case, the braking valve can comprise a braking valve piston, which is movable in a braking valve housing, cooperates with a control edge of the braking valve housing and has a bevel. The braking valve piston can strike against an adjustable stop which defines the throttle area of the braking valve, which throttle area is fixed by the overlap of the bevel of the braking valve piston with the control edge of the braking valve housing. In said case, the braking valve can comprise a braking valve spring which biases the braking valve piston towards the stop. 
     The control valve can take the form of seat valves and each comprise a control valve piston, which is movable in each case in a control valve housing. In said case, the control valve piston can have a conical portion, which cooperates with a valve seat so as to form a sealed seat. It is advantageous for the control valves to taken the form of seat valves because they then present a relatively high pressure resistance and insensitivity to fouling. Each control valve can comprise a control valve spring, which pressure the control valve piston against the valve seat. The control valve piston preferably takes the form of a stepped piston, wherein a step of the control valve piston is biased by the activating control pressure, thereby producing a hydraulically activated seat valve. 
     The braking valves and the control valves can be connected by a pressure change valve to the control lines. A supply device can be provided, which generates a supply pressure in a supply line. The actuating pressure can be connected in each case by an associated pressure control valve to the supply line, wherein the actuating pressure in the actuating pressure lines is adjusted by means of the control pressure prevailing in the control lines. When a pressure control valve spring is provided, which sets the actuating pressure slightly higher than the activating control pressure, then even given an imperceptible control pressure, there is a slight actuating pressure available, which is used to top up the actuating pressure chamber which increases in volume when the hydraulic pump swings back. A top-up device with a relatively large filter is therefore not required. 
     The control lines can be alternately loadable with control pressure by means of a control transmitter, which is connected to a control pressure supply and the hydraulic fluid tank. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     There now follows a description of preferred embodiments with reference to the drawings. The drawings show: 
     FIG. 1 a hydraulic block diagram of a first embodiment of the hydraulic control system according to the invention; 
     FIG. 2 a hydraulic block diagram of a second embodiment of the hydraulic control system according to the invention; and 
     FIG. 3 a diagrammatic constructional realization of the embodiment shown in FIG.  1 . 
    
    
     DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS 
     FIG. 1 shows a first embodiment of the hydraulic control system according to the invention. The hydraulic control system denoted generally by the reference character  1  is used in particular to activate the slewing gear of a digger. The slewing gear of the digger is in said case driven by a hydraulic motor (not shown), which is connected by a first working line  2  and a second working line  3  to the hydraulic pump  4  so as to form a working circuit. The hydraulic pump  4 , e.g. for an i.c. engine (not shown), is driven via the drive shaft  5 . The delivery direction of the hydraulic pump is reversible so that, depending on the desired direction of rotation of the slewing gear, either the working line  2  or the working line  3  operates as a high-pressure line. 
     The displacement volume of the hydraulic pump  4  is adjustable by means of an adjusting apparatus  6 . The adjusting apparatus  6  comprises an actuating piston  7 , which is movable in an actuating cylinder  8  and is centered without pressurization in its neutral position with zero displacement volume, which is shown in FIG. 1, by means of two centering springs  9  and  10 . The actuating piston  7  divides the actuating cylinder  8  into a first actuating pressure chamber  11  and a second actuating pressure chamber  12 . The first actuating pressure chamber  11  is connected to a first actuating pressure line  13 , while the second actuating pressure chamber  12  is connected to a second actuating pressure line  14 , which lines supply the actuating pressure to the actuating pressure chambers  11 , 12 . 
     According to the invention, a branch  15 ,  16  is provided in each of the actuating pressure lines  13 ,  14 . A secondary line  17 ,  18  branches off towards each braking valve  19 ,  20  so that the first actuating pressure chamber  11  is connected by the braking valve  19  to the hydraulic fluid tank  21  and the second actuating pressure chamber  12  is connected by the braking valve  20  to the hydraulic fluid tank  21 . The braking valve  19 ,  20  has a closed valve position  22 ,  23 , in which the flow through the respective braking valve  19 ,  20  is interrupted, and a throttled valve position  24 ,  25 , in which the flow through the respective braking valve  19 ,  20  is throttled. The throttle area, which the braking valve  19 ,  20  has in its throttled valve position  24 ,  25 , is preferably adjustable. The braking valves  19  and  20  are activated by a common control pressure line  26  in such a way that, when the control pressure in the control pressure line  26  drops below a defined threshold value, they change or switch over into their throttled valve position  24  and  25  respectively. When the control pressure in the control pressure line  26  exceeds the defined threshold value, the braking valves  19  and  20  are situated in their closed valve position  22  and  23  respectively and are blocked. When however the control pressure in the control pressure line  26  is greater than the defined threshold value, the braking valves  19  and  20  are pressed into their throttled valve position  24  and  25  respectively so that the braking valves  19  and  20  have a throttled, preferably adjustable throughflow. The threshold value is preferably preset to a very low, almost or totally imperceptible control pressure and is adjustable by means of the braking valve springs  29  and  30 . 
     Situated in each of the actuating pressure lines  13  and  14  is a control valve  27  and  28  respectively. Said control valves  27  and  28  are disposed in such a way that the branches  15  and  16  are situated in each case between the control valves  27  and  28  and the actuating pressure chambers  11  and  12  of the adjusting apparatus  6 . The braking valves  19  and  20  are therefore connected by the branches  15  and  16  directly to their associated actuating pressure chamber  11  and  12  respectively without any further hydraulic valves, besides the braking valves  19  and  20 , being situated along the hydraulic line path between the actuating pressure chambers  11  and  12  and the hydraulic fluid tank  21 . The braking valves  19  and  20  are preferably disposed in the immediate spatial vicinity of the actuating pressure chambers  11  and  12 , using only short line paths for the line portion of the actuating pressure line  13 ,  14  to the branch  15 ,  16  and for the secondary line  17 ,  18 . 
     The control valves  27  and  28  are activated likewise by the control pressure prevailing in the control pressure line  26 . In said case, the control valves  27  and  28  open when the control pressure in the control pressure line  26  exceeds a defined threshold value. Conversely, the control valves  27  and  28  close when the control pressure in the control pressure line  26  drops below the defined threshold value. The control valves  27  and  28  preferably take the form of seat valves, e.g. check valves, while the braking valves  19  and  20  preferably take the form of slide valves. 
     In the illustrated embodiment, the actuating pressure in the actuating pressure lines  13  and  14  and hence the deflection of the hydraulic pump  4  is defined by means of a manual control transmitter  32 , which connects two control lines  33  and  34  either to a control pressure supply  35  or to the hydraulic fluid tank  21  depending on the desired direction of rotation of the slewing gear. Depending on the intended direction of rotation of the slewing gear, either the control line  33  or the control line  34  is loaded with control pressure. In the embodiment, the control lines  33  and  34  are directly connected by throttle points  36  and  37  to the control valves  27  and  28 . In the embodiment illustrated in FIG. 1, the actuating pressure prevailing in the actuating pressure lines  13  and  14  is therefore derived directly from the control pressures prevailing in the control lines  33  and  34 . Said embodiment dispenses with pilot control and is suitable particularly for slewing gear control systems of a small nominal size. 
     The control lines  33  and  34  are connected to the control pressure line  26  by a pressure change valve  38 , which in each case selects the highest of the control pressures prevailing in the two control lines  33  and  34 . In each case, the highest of the control pressures prevailing in the control lines  33  and  34  therefore prevails in the control pressure line  26 . The control pressure line  26  is connected by a pressure cut-off valve  39  to the hydraulic fluid tank  21 . The pressure cut-off valve  39  takes the form of a pressure relief valve and limits the pressure in the control pressure line  26  to a maximum pressure defined by means of an electrical transmitter  40 . The control pressure line  26  is connected to the hydraulic fluid tank  21  by a further pressure relief valve  41 , which is activated via a pressure change valve  42  by the, in each case, highest working pressure prevailing in the working lines  2  and  3  and enables working pressure-dependent pressure relief. 
     A supply device  43  is further provided. The supply device  43  comprises a supply pump  44 , which is connected by the common shaft  5  to the hydraulic pump  4  and via a supply filter  45  supplies a supply pressure limited by the pressure relief valve  47  into a supply line  46 . The supply pressure is introduced via a check valve  48  or  49  into the respective working line  2  or  3  carrying the low pressure. The maximum working pressure in the working lines  2  and  3  is limited by the pressure relief valves  50  and  51 . 
     The hydraulic control system according to the invention operates in the following way. 
     To accelerate the slewing gear driven by the hydraulic motor (not shown), the hydraulic pump  4  connected to the hydraulic motor is swung out by operating the joy-stick  53  of the control transmitter  32 . Depending on the intended direction of rotation of the slewing gear, either the control line  33  or the control line  34  is loaded via the control pressure supply  35  with a proportioned control pressure, while the other control line  34  or  33  is connected to the hydraulic fluid tank  21 . The control pressure building up in the control line  33  or prevails also in the control pressure line  26  and effects an opening of the control valves  27  and  28 . In the embodiment illustrated in FIG. 1, the actuating pressure lines  13  and  14  are therefore connected by the control valves  27  and  28  directly from the control lines  33  and  34 , so that the actuating pressure in the illustrated embodiment is derived directly to the control pressure. As a result, one of the two actuating pressure chambers  11  or  12  is loaded with actuating pressure and the other actuating pressure chamber  12  or  11  is relieved via the respective control valve  27  or  28  and the control transmitter  32  towards the hydraulic fluid tank  21 . The actuating piston  7  of the adjusting apparatus  6  is accordingly displaced and the hydraulic pump  4  is swung out in the intended direction. The braking valves  19  and  21  are biased by the control pressure in the control pressure line  26  in such a way that they are situated in their closed valve position  22  and  23  and so via the braking valves  19  and  20  no pressure losses arise in the actuating pressure lines  13  and  14 . 
     As soon as the slewing gear has reached the desired rotational speed, the operator may let go of the joy-stick  53  with the result that the control transmitter  32  is returned into its neutral position, in which it connects the control lines  33  and  34  to the hydraulic fluid tank  21 . Thus, control pressure no longer prevails in the control lines  33  and  34  and the common control pressure line  26  also no longer carries control pressure. Consequently, the control valves  27  and  28  are closed by the control valve spring  54  and  55 , while the braking valves  19  and  20  are switched by their braking valve springs  29  and  30  into their throttled valve position  24  and  25 . The hydraulic pump  4  is still situated in its swung-out delivery position with the actuating piston  7  displaced out of the neutral position. The centring springs  9  and  10  gradually return the actuating piston  7  into its neutral position shown in FIG. 1, wherein the time constant required for said purpose depends upon the throttling effected by the braking valves  19  and  20 . Since the throttling of the return flow of hydraulic fluid from the actuating pressure chambers  11  and  12  to the hydraulic fluid tank  21  is determined almost exclusively by the throttle area of the respective braking valve  19  or  20 , said time constant may be adjusted very precisely and reproducibly. Since the throttle area of the braking valves  19  and  20  is preferably designed so as to be variable, a suitable fine tuning may be effected. According to the invention, the braking valves  19  and  20  are connected directly, without interposing further valves or longer hydraulic lines, to the actuating pressure chambers  11  and  12  with the result that the effective throttling of the return flow is determined solely by the braking valves  19  and  20 . A reflux of hydraulic fluid into the control lines  33  and  34  is ruled out because the control valves  27  and  28  block in said operating situation. 
     The threshold value for the switchover between the valve positions of the braking valves  19  and  20  and the control valves  27  and  28  is adjustable by means of the braking valve springs  29  and  30  and the control valve springs  54  and  55  respectively. 
     FIG. 2 shows a second embodiment of the hydraulic control system according to the invention. Elements already described with reference to FIG. 1 are provided with matching reference characters so that, in said respect, a repeat description is unnecessary. 
     The embodiment shown in FIG. 2 differs from the embodiment already described with reference to FIG. 1 in that two pressure control valves  60  and  61  are provided, which at their outputs are connected to the actuating pressure lines  13  and  14  in each case upstream of the control valves  27  and  28 . A respective one of the inputs of the pressure control valves  60  and  61  is connected to the hydraulic fluid tank  21 , while a respective other input of the pressure control valves  60  and  61  is connected by a connecting line  62  in each case to the supply line  46 . Each pressure control valve  60  or  61  is connected at a first control input to an associated control line  33  or  34  and at a second control input to the actuating pressure line  13  or  14  by a detour line  63  or  64 . Each pressure control valve  60  or  61  is therefore activated by a pressure difference between the control pressure in the associated control line  33  or  34  and the actuating pressure in the associated actuating pressure line  13  or  14 . As a result, the actuating pressure in the actuating pressure line  13  or  14  substantially corresponds to the control pressure in the associated control line  33  or  34 . 
     Since the pressure control valves  60  and  61  are in addition biased slightly in the opening direction by a pressure control valve spring  66  and  67  respectively, the actuating pressure prevailing in the actuating pressure line  13  or  14  is slightly, e.g. 1 to 2 bar, higher than the control pressure in the associated control line  33  or  34 . In the actuating pressure line a slight pressure therefore prevails even when there is no control pressure in the associated control line  33  or  34 . During the return of the actuating piston  7  into its neutral position defined by the centering springs  9  and  10 , hydraulic fluid may therefore continue to flow via the supply device  43 , the connecting line  62  and the associated pressure control valve  60  or  61  as well as the associated control valve  27  or  28  into the actuating pressure chamber  11  or  12  which increases in volume during the return of the actuating piston  7  into the neutral position. A top-up device with a correspondingly large top-up filter is therefore not required. 
     By virtue of the reduction of the control pressure-dependent actuating pressure effected by means of the pressure control valves  60  and  61 , the embodiment shown in FIG. 2 is also suitable for hydraulic control systems of a large nominal size, i.e. for large-dimension slewing gear control systems. 
     FIG. 3 shows a diagrammatic view of an exemplary constructional refinement of the braking valves  19  and  20  and the control valves  27  and  28 . To make it easier to understand, the hydraulic circuit in accordance with FIG. 1 is likewise indicated. Elements already described with reference to FIG. 1 are provided with matching reference characters so that, in said respect, a repeat description is unnecessary. 
     In the preferred embodiment shown in FIG. 3, the braking valves  19  and  20  take the form of slide valves. Braking valve pistons  80  and  81  are in each case disposed in an axially movable manner in a braking valve housing  82  or  83  and biased by means of the braking valve spring  29  or  30  towards a preferably adjustable stop  84  or  85 . The stop  84  or  85  projects axially in a cylinder bore  86  or  87  formed in the respective braking valve housing  82  or  83 . The extent of axial projection may be adjusted, for example, in that the stop  84  or  85  has a thread which may be screwed into the braking valve housing  82  or  83 . The position of the stops  84  and  85  may alternatively be adjustable by means of an e.g. electromagnetic or hydraulic transmitter by the operator of the digger so that the slow, gentle outward swing of the slewing gear may be flexibly adjusted by varying the throttle area of the braking valves  19  and  20  by means of the stops  84  and  85 . 
     The braking valve piston  80  or  81  has a bevel  88  or  89  and cooperates with a control edge  92  or  93  formed on an annular groove  90  or  91 . The control pressure line  26  leads to a pressure chamber  94  or  95 , to which the braking valve piston  80  or  81  is adjacent. As the pressure in the control pressure line  26  increases, the braking valve piston  80  or  81  is therefore displaced towards the braking valve spring  29  or  30  and the control edge  92  or  93  is sealed by the non-bevelled region of the braking valve piston  80  or  81 . As the pressure in the control pressure line  26  decreases, the braking valve piston  80  or  81  is retracted in FIG. 3 to the left or right by the braking valve spring  29  or  30  so that the bevel  88  or  89  progressively releases the control edge  92  or  93 . The throttle opening of the braking valve  19  or  20  in the position of abutment against the stop  84  or  85  is fixed by the position of the stop  84  or  85  and is adjustable by varying the position of the stop  84  or  85 . 
     In the preferred embodiment shown in FIG. 3, the control valves  27  and  28  take the form of seat valves. The control valve pistons  96  and  97  are movable in each case in a control valve housing  98  or  99 . The control valve pistons  96  and  97  each have a conical portion  100  or  101 . The control valve pistons  96  and  97  are each biased by a control valve spring  54  and  55  in such a way that the conical portion  100  or  101  is pressed against the valve seat  102  or  103  so as to produce a sealed seat. Formed upstream of the conical portion  100  or  101  is a first valve chamber  104  or  105 , which is connected to the valve input. In the embodiment shown in FIG. 3, the valve input is connected directly to the associated control line  33  or  34 . The valve output is connected to the associated actuating pressure line  13  or  14 . In each case, a second valve chamber  106  or  107  is isolated from the first valve chamber  104  or  105  by a sealing step  108  or  109  of the control valve piston  96  or  97  and connected to the control pressure line  26 . The control pressure prevailing in the control pressure line  26  acts upon a surface  110  or  111  of the control valve piston  96  or  97  and displaces the control valve piston  96  or  97  towards the control valve spring  54  or  55 . When threshold valve defined by the control valve spring  54   55  is exceeded, the conical portion  100  or  101  lifts off the valve seat  102  or  103  and enables the flow through the control valve  27  or  28 . 
     The braking valves  19  and  20  and the seat valves  27  and  28  may alternatively be designed in a different manner. In particular, it is possible for the control valves  27  and  28  to be alternatively designed as simple check valves, which prevent a reflux of hydraulic fluid into the control line  33  and  34  and/or into the pressure control valves  60  and  61 .