Abstract:
A pressure reservoir is used to exert pressure on a hydraulic system with which, a gas exchange valve, for instance, of an internal combustion engine can be actuated. The pressure reservoir includes a housing and a piston that is prestressed in operation by a device. To enable making the pressure reservoir as small as possible, it is proposed that the device which prestresses the piston of the pressure reservoir has a characteristic force-travel curve, in one range of motion of the piston, that has a slope which differs from the slope in a different range of motion of the piston.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a 35 USC 371 application of PCT/DE 02/00079, filed on Jan. 12, 2002. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to a pressure reservoir for exerting pressure on a hydraulic system, with which preferably a gas exchange valve of an internal combustion engine is actuated, having a housing and a piston prestressed in operation by a device. 
     2. Description of the Prior Art 
     A hydraulic system with a pressure reservoir of the type with which this invention is concerned is known from German Patent Disclosure DE 193 26 047 A1. A hydraulic system of this kind is used for instance for actuating the inlet and outlet valves of an internal combustion engine, if the engine does not have a camshaft. Such an engine has the advantage that the control times of the inlet and outlet valves are independent of the position of the piston of the applicable cylinder. Depending on the engine operating state, such as high rpm, and on the torque desired by the driver, valve opening and closing times can be achieved which make especially optimal engine operation possible in terms of emissions and fuel consumption. 
     The known hydraulic system functions with a hydraulic circuit, which is supplied from a hydraulic reservoir via a high-pressure hydraulic pump. An actuating device includes a piston that can be acted upon hydraulically in both directions of motion and that is connected to the valve shaft of a gas exchange valve, such as an inlet valve. Via 2/2-way valves, one at a time of the two chambers of the hydraulic cylinder can be subjected to higher pressure, which leads to a corresponding motion of the piston and as a result to an opening or closing event of the gas exchange valve of the engine block. 
     The hydraulic circuit communicates with a hydraulic pressure reservoir, which is embodied as a spring-loaded piston reservoir and serves to damp vibration in the hydraulic system. An identically embodied emergency pressure reservoir also communicates with one of the two chambers in the hydraulic cylinder; if the pressure drops in the hydraulic line, this emergency pressure reservoir still furnishes sufficient pressure and a sufficient fluid volume to enable the gas exchange valve to be moved to its closed position of repose. The two pressure reservoirs operate at different pressure levels, which are set by means of different stiffnesses of the corresponding restoring springs. From DE 198 26 047 A1, it is also known to use only a single pressure reservoir, which functions simultaneously as both a working pressure reservoir and an emergency pressure reservoir. 
     If only a single pressure reservoir is provided, its design must be such that at minimal operating pressure in the hydraulic system, sufficient hydraulic medium is stored to enable reliably moving the gas exchange valve into the closed position of repose in the event of an emergency. This requires a relatively soft spring and a long spring travel. In order at the same time to assure that over the entire operating pressure range, a sufficient damping action exists, this kind of pressure reservoir, equipped with a soft spring, must be very long structurally, as a function of the minimum and maximum operating pressure. Such a large pressure reservoir, however, can be accommodated only with difficulty in the available installation space in an internal combustion engine. Moreover, because of the great structural length, in the operating pressure range a relatively large volume of fluid must be stored in such a pressure reservoir, and as an idle volume, beyond the desired damping action, this adversely affects the dynamics of the hydraulic system. 
     It is therefore the object of the present invention to refine a pressure reservoir of the type defined at the outset such that on the one hand, a pressure damping function and on the other an emergency pressure function are available, while nevertheless the pressure reservoir is as small as possible. 
     The above and other objects and advantages are attained, in a pressure reservoir of the type defined at the outset, by providing that the device which prestresses the piston of the pressure reservoir has a characteristic force-travel curve, in one range of motion of the piston, that has a slope which differs from the slope in a different range of motion of the piston. 
     According to the invention, a prestressing device with a nonlinear characteristic is used in the pressure reservoir. It is understood then that first, when the piston is urged out of its pressureless position of repose, a softer characteristic of the prestressing device is desired; that is, a change in pressure results in a relatively long movement distance of the piston. In a range of motion of the piston that is far away from the position of repose of the piston, conversely, a stiffer characteristic of the prestressing device of the piston is desired; that is, a pressure change should cause only a comparatively slight motion of the piston. 
     In this way, both desired functions, namely the emergency pressure function and the vibration damping function, can be achieved in a single pressure reservoir: The emergency pressure function is available in the range of motion of the piston of the pressure reservoir in which the prestressing device has a relatively soft characteristic. Within this piston range of motion, the pressure reservoir is thus already capable, at only a slight pressure drop, of dispensing a large enough fluid volume into the hydraulic circuit for securing, for instance a gas exchange valve, in the event of a pressure loss. The vibration damping function exists in the range of motion of the piston within which the characteristic force-travel curve is comparatively steep. In this piston range of motion, even major pressure fluctuations result in only a slight piston motion. Accordingly, in this piston range of motion, it is also possible for only a slight movement distance of the prestressing device to be provided, which in turn is favorable for the sake of a short structural length of the pressure reservoir. 
     The pressure reservoir of the invention can accordingly be used on the one hand for storing a fluid volume for emergency operation, and on the other, it can be used in normal operation for vibration damping, and at the same time is very small in size. It can therefore be integrated easily and without problems into the available installation space. Furthermore, because of the slight fluid volume stored and the great stiffness of the prestressing device, an optimal vibration damping can be achieved in normal operation without impairing the system dynamics. 
     In a first refinement, the device which prestresses the piston of the pressure reservoir has at least two series-connected devices, with characteristic force-travel curves of different slope, which prestress the piston in operation. The desired properties of such a pressure reservoir can be achieved especially easily, since in it, the various functions are also performed physically separately. 
     It is especially preferred that the devices for prestressing the piston include at least two series-connected springs, and the stiffness of one spring differs from that of the other spring. A pressure reservoir with this kind of two-stage spring assembly can be constructed simply and very economically and furthermore is robust. 
     In an especially preferred feature of the pressure reservoir of the invention, the pressure reservoir has an elongated part with two end portions and one support portion, which is disposed between the end portions and has a larger outer dimension than the end portions and on which two adjacent springs are braced, the one spring being tightened in operation between one side of the support portion and the piston, and the other spring being tightened between the other side of the support portion and a housing portion. An elongated part of this kind enables the secure guidance of the piston, on the one hand, and of the corresponding springs, on the other. 
     It is also provided that at least two stops are provided, which prevent the springs from being tightened into a block in operation. Essentially, tightening springs into a block has two disadvantages: First, most springs, in the range of motion located just before tightening into a block occurs, exhibit a markedly nonlinear, and above all often non-replicable, characteristic curve behavior. This is unwanted in the present case as well. Furthermore, whenever the springs are tightened into a block, wear of the touching surfaces of the springs can occur, which can impair the service life of the springs. The stops according to the invention prevent this. 
     Especially simply, such stops can be realized in conjunction with the above-described elongated part: In this case, the length of the elongated part can be adapted such that one axial end of the elongated part forms a stop with a housing portion of the pressure reservoir, and the other axial end of the elongated part forms a stop with the piston. 
     Basically, all types of springs are suitable for the pressure reservoir of the invention. Examples are spiral springs, air springs and magnet springs. It is especially preferred, however, that at least one of the springs is a cup spring. The use of cup springs, because of the better ratio between the spring work and the installation space, brings about a further reduction in the structural length of the pressure reservoir. Moreover, because of the strong friction damping in a cup spring assembly, the damping action of the reservoir is enhanced. 
     The invention also relates to a hydraulic system for actuating a gas exchange valve of an internal combustion engine, in particular of a motor vehicle, having a fluid reservoir, a fluid pump, a fluid line, a pressure reservoir that communicates with the fluid line having a housing and a piston prestressed in operation by a device, and having an actuating device, which communicates via a valve device with the fluid line and actuates the gas exchange valve. 
     To reduce the overall dimensions of the hydraulic system, it is proposed that the pressure reservoir be embodied as described above. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     Below, exemplary embodiments of the invention are described in detail, in conjunction with the accompanying drawings, in which: 
     FIG. 1, a basic illustration of a hydraulic system for actuating a gas exchange valve of an internal combustion engine; 
     FIG. 2, a section through a first exemplary embodiment of a pressure reservoir of the hydraulic system of FIG. 1; 
     FIG. 3, a pressure and travel graph to explain the function of the pressure reservoir of FIG. 2; 
     FIG. 4, a schematic section through a second exemplary embodiment of a pressure reservoir; 
     FIG. 5, a schematic section through a third exemplary embodiment of a pressure reservoir; 
     FIG. 6, a schematic section through a fourth exemplary embodiment of a pressure reservoir; and 
     FIG. 7, a schematic section through a fifth exemplary embodiment of a pressure reservoir. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     In FIG. 1, a hydraulic system referred to overall by reference numeral  10  serves to actuate a gas exchange valve, which here is embodied as an inlet valve  12  of an internal combustion engine  14 . 
     The inlet valve  12  is actuated by a hydraulic cylinder  16 . This cylinder includes a housing  18 , in which a piston  20  with a piston rod  22  is guided slidingly. The piston rod  22  is passed through the housing  18  and is connected to a valve shaft  24 , which in turn is formed onto a platelike valve element  26 . In the closed state of the inlet valve  12 , the valve element  26  rests tightly against a valve seat  28  in the upper region of a combustion chamber  30  of the engine  14 . If no hydraulic pressure is available, the piston  20  is pressed upward by a spring  32 , and as a result the inlet valve  12  is closed. 
     The hydraulic system  10  further includes a supply container  34 , from which hydraulic fluid is pumped by a high-pressure pump  36  into a high-pressure hydraulic line  38 . Downstream of a check valve  40 , the high-pressure hydraulic line  38  branches off into one branch  42 , which discharges directly into a lower work chamber  44  of the hydraulic cylinder  16 . Another branch  46  of the high-pressure hydraulic line  38  leads to a 2/2-way switching valve  48 , which in the currentless state is pressed into its closed position by a spring  50 . The branch  46  of the high-pressure hydraulic line  38  leads, downstream of the 2/2-way switching valve  48 , to an upper work chamber  52  of the hydraulic cylinder  16 . From there, a high-pressure hydraulic line leads, via a further 2/2-way switching valve  56  and a check valve  58 , back to the supply container  34 . The 2/2-way switching valve  56  is opened by a spring  57 , in the currentless state. 
     A tie line  60 , which communicates with a pressure reservoir  62 , discharges at the point where the high-pressure hydraulic line  38  branches off into the branch  42  and the branch  46 . The construction of the pressure reservoir is shown in detail in FIG.  2 . 
     The pressure reservoir  62  includes a housing  64 , which has an overall cylindrical shape, and in which a cylindrical hollow chamber  66  is embodied. On the right-hand side, in FIG. 2, the hollow chamber  66  is closed with a cap  68 , while conversely, on the left-hand side in FIG. 2, it communicates with the tie line  60  via a connecting conduit  70 . The cap  68  has a valve opening, which in the present exemplary embodiment is located outside the sectional plane and is therefore not visible. 
     A piston  72  is retained displaceably in the hollow chamber  66 . The radial jacket face of the piston  72  is sealed off from the inner wall of the hollow chamber  66  by a sealing ring  74 , which is placed in an annular groove  76  in the outer jacket face of the piston  72 . A piston rod  78  is formed onto the piston  72 . It extends from the piston  72  toward the cap  68 . The piston  72  and the piston rod  78  are coaxial to the hollow chamber  66  of the housing  64  of the pressure reservoir  62 . 
     Coaxially to the piston  72  and to the piston rod  78 , there is an elongated tubular part  80  located in the hollow chamber  66  of the pressure reservoir  62 . The elongated tubular part  80  is slipped onto the piston rod  78  in sliding communication. The elongated tubular part  80  includes a cylindrical end portion  82 , located on its left-hand side in terms of FIG. 2, and a cylindrical end portion  84 , located on its right-hand side in FIG.  2 . Located between the two end portions  82  and  84  is a support portion  86 , whose outside diameter is greater than the outside diameter of the left-hand end portion  82  and of the right-hand end portion  84 . In other words, the support portion  86  takes the form of an annular collar. 
     Between the support portion  86  and the piston  72 , a packet  87  of a total of twelve cup springs  88  (for the sake of simplicity, not all the cup springs  88  have reference numerals in the drawing) is disposed coaxially to the piston  72 , piston rod  78 , and elongated tubular part  80 . The packet  87  is divided into four individual groups (not carrying reference numerals), each comprising three parallel cup springs  88 . A packet  89  comprising three parallel cup springs  90  is disposed between the support portion  86  and the cap  68  of the housing  64 . 
     In the pressureless state of repose, shown in FIG. 2, of the pressure reservoir  62 , the cup springs  88  and  90  are relaxed. In this state, there is a free space between the axial end, on the left in FIG. 2, of the elongated tubular part  80  and the piston  72 . A free space is also present between the right-hand axial end, in the drawing, of the elongated tubular part  80  and the bottom of a recess  92  in the cap  68  of the housing  64 . The cup springs  88  are all softer than the cup springs  90 . The spring travel of the packet formed of the cup springs  88  is overall longer than the spring travel of the group formed by the cup springs  90 . 
     The hydraulic system  10  shown in FIG. 1, having the pressure reservoir  62  shown in FIG. 2, functions as follows: 
     The high-pressure pump  36  pumps hydraulic fluid out of the supply container  34  into the hydraulic line  38  and from there via the branch line  42  into the lower work chamber  44  of the hydraulic cylinder  16 . When the switching valve  48  is opened and the switching valve  56  is closed, the upper work chamber  52  of the hydraulic cylinder  16  is also put under pressure by hydraulic fluid. Since the engagement area in the axial direction on the top side of the piston  20  of the hydraulic cylinder  16  is greater than on its underside, the piston  20  is pressed downward in this case, and the inlet valve  12  is opened. 
     If the switching valve  48  is closed and the switching valve  56  is opened, the upper work chamber  52  is made to communicate, via the branch line  54 , with the ambient pressure, and as a result the piston  20  is moved upward again, and the inlet valve  12  is closed. In this way, without having to trigger the inlet valve  12  mechanically, for instance by means of a camshaft of the engine  14 , very fast opening and closing times of the inlet valve  12  can be attained. 
     If the high-pressure pump  36  is not pumping, and in other words the hydraulic line  38  and the tie line  60  are pressureless, then the piston  72  of the pressure reservoir  62  is in the position of repose shown in FIG.  2 . In the graph of FIG. 3, in which the travel s of the piston  72  of the pressure reservoir  62  is plotted over the hydraulic pressure p, this position of repose is at a position identified by reference numeral  94 . 
     If the high-pressure pump  36  is switched on, the pressure in the hydraulic line  38  and the tie line  60  rises. Since the cup springs  88  have a lesser stiffness than the cup springs  90 , the elongated tubular part  80  initially remains stationary during this pressure increase, while conversely the piston  72  moves in the direction of the cap  68  of the housing  64  and in the process compresses the cup springs  88 . 
     The spacing between the left-hand axial end, in terms of FIG. 2, of the elongated tubular part  80  and the piston  72  is selected such that the piston  72  comes to rest on the elongated tubular part  80  whenever the minimum operating pressure PBMIN is reached. The corresponding travel accomplished by the piston  72  is shown in FIG. 3 as SPBMIN. The geometry inside the pressure reservoir  62 , and in particular the length of the left-hand end portion  82  of the elongated tubular part  80 , is selected such that whenever the piston  72  comes to rest on the elongated tubular part  80 , the cup springs  88  have not yet moved into a block. 
     If the pressure is increased further, then the elongated tubular part  80  is moved by the piston  72  in the direction of the bottom of the recess  92  in the cap  68  of the housing  64 . As a result, the cup springs  90  are deformed. Since the cup springs  90  are considerably stiffer than the cup springs  88 , in this range a markedly greater slope of the curve shown in FIG. 3 results. The spacing between the right-hand axial end, in terms of FIG. 2, of the elongated tubular part  80  and the bottom of the recess  92  in the cap  68  is selected such that whenever the hydraulic pressure reaches the maximum operating pressure PBMAX, the elongated tubular part  80  comes to rest on the bottom of the recess  92  in the cap  68 . The length of the right-hand end portion  84  of the elongated tubular part  80 , in turn, is selected such that whenever the elongated tubular part  80  touches the cap  68 , the springs  90  of the group  89  have not yet been completely deformed. The piston  72  in this case has covered the maximum possible travel SPBMAX. 
     When the hydraulic system  10  is in its normal operating state, the hydraulic pressure in the hydraulic lines  38 ,  42 ,  46  and  60  is in the range between the minimum operating pressure PBMIN and the maximum operating pressure PBMAX. In this case, the pressure reservoir  62  functions as a vibration damper for pressure fluctuations that occur in the hydraulic fluid of the hydraulic system  10 . Because of the great stiffness of the cup springs  90 , even major amplitudes of the pressure vibrations cause only a slight motion of the piston  72 . The length of the packet  89  of cup springs  90  can therefore be slight, which in turn reduces the total structural length of the pressure reservoir  62 . 
     The great stiffness of the cup springs  90  also makes it possible to reduce the fluid volume stored in the pressure reservoir  62 . This makes the desired vibration damping in the operating pressure range possible, without impairment of the system dynamics of the hydraulic system  10 . Moreover, the use of the cup springs  90  improves the damping action of the pressure reservoir  62 , since major friction damping occurs between the individual cup springs  90 . 
     Compared to a conventional pressure reservoir, the pressure reservoir  62  shown in FIG. 2 is very small in size. If vibration damping in the same operating pressure range is to be furnished in a conventional pressure reservoir, the conventional pressure reservoir would have to have a markedly longer spring travel and thus a markedly greater structural length. This is represented by dashed lines in FIG.  3 . The spring travel required in a conventional pressure reservoir for the same operating pressure range and the same emergency pressure properties is marked SPMAX′ in FIG.  3 . The reduction in structural length for the pressure reservoir  62  compared with a conventional pressure reservoir thus amounts to the difference between SPMAX′ and SPMAX. 
     Upon a pressure drop inside the hydraulic system  10 , for caused by a failure of the high-pressure pump  36 , assurance must be provided that the piston  20  of the hydraulic cylinder  16  can still be moved far enough upward that the inlet valve  12  can be closed. This is necessary to prevent the valve element  26  of the inlet valve  12 , which element protrudes into the combustion chamber  30 , from colliding with other valve elements or even with the piston (not shown) in the combustion chamber  30 . 
     In such a case, the cup springs  90  and especially the cup springs  88  press the piston  72  in the pressure reservoir  62  back into its extreme left-hand position in FIG.  2 . Correspondingly, a hydraulic fluid volume is forced out of the pressure reservoir  62  into the tie line  60  and from there via the branch line  42  into the lower work chamber  44  of the hydraulic cylinder  16 . The spring travel of the cup springs  88  and the resultant movement distance SPMIN of the piston  72  is selected such that secure closure of the inlet valve  12  is possible in every situation. Thus in the normal operating range, a pressure reservoir  62  with optimal damping properties is available, while conversely, in the event of a pressure drop, the same pressure reservoir  62  furnishes a sufficient hydraulic fluid volume for secure closure of the inlet valve  12  via the hydraulic cylinder  16 . 
     In FIGS. 4-7, further exemplary embodiments of pressure reservoirs  62  are shown schematically. Elements whose function is equivalent to those shown in FIG. 2 are identified by the same reference numerals. They will not be described again in detail. 
     In the exemplary embodiment shown in FIG. 4, an elongated tubular part  80  is omitted. Instead, the springs  88  and  90 , shown only symbolically, of different stiffness and different length are integrally joined together. 
     In the exemplary embodiment shown in FIG. 5, instead of cup springs or helical springs, air springs  88  and  90  are used, which have different volumes and different fill pressures. 
     In FIG. 6, springs of equal stiffness are used, but these are springs disposed parallel, with different lengths. The spring  88  disposed centrally in FIG. 6 has a greater length than the two springs  90  disposed laterally of the spring  88 . In this way, in a first range of motion of the piston  72 , located adjacent to the position repose, only the spring  88  is initially acted upon, while conversely in a second range of motion of the piston  72 , the springs  90  are acted upon as well, as a result of which the total spring stiffness increases. 
     In FIG. 7, instead of springs, an electromagnet  88  is used, which exerts a repellent force on the piston  72  made of a permanent magnetic material. The repellent force can be adjusted by means of a controller  96  as a function of the position of the piston  72 , which position is detected by a sensor  98 . 
     The foregoing relates to preferred exemplary embodiments of the invention, it being understood that other variants and embodiments thereof are possible within the spirit and scope of the invention, the latter being defined by the appended claims.