Abstract:
The invention relates to an electrohydraulic valve controller for controlling a gas exchange valve in internal combustion engines, having a hydraulically actuatable control valve including a control valve piston which can be acted upon, via electrically actuatable valves, by a hydraulic medium that is under pressure, and a hydraulically acting valve brake is assigned to the control valve piston. The valve brake ( 46 ) includes a temperature compensation for the hydraulic medium.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This application is a 35 USC 371 application of PCT/DE 02/01806 filed on May 18, 2002. 

   BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The invention relates to an improved electrohydraulic valve controller, in particular for controlling a gas exchange valve in internal combustion engines. 
   2. Prior Art 
   In internal combustion engines used for driving motor vehicles a fuel-air mixture is compressed and ignited in a work chamber where the energy produced is converted into mechanical work. In such engines, the air, or the fuel-air mixture, is delivered to the work chamber via valves (inlet or intake valves) and to remove the products of combustion from the work chamber via valves (outlet or exhaust valves). Controlling these valves is very significant for determining the efficiency of the engine. In particular, the gas exchange in the work chamber is controlled via the control of the valves. 
   Besides camshaft control, it is also known to use an electrohydraulic valve controller. The electrohydraulic valve controller offers the capability of variable or fully variable valve control, making it possible to optimize the gas exchange and thus to enhance the motor efficiency [redundant, or “engine efficiency”] of the engine. 
   The electrohydraulic valve controller includes a hydraulically actuatable control valve, whose control valve piston actuates a valve body of the inlet and outlet valves and leads toward a valve seat (valve seat ring) (closure of the valve) or moves away from it (opening of the valve). The control valve can be actuated by way of controlling the pressure of a hydraulic medium. The pressure control is effected here via magnet valves incorporated into the hydraulic circuit. To achieve gas exchanges that are as optimal as possible, the highest possible switching speeds of the control valve are needed. As a result of these high switching speeds, the valve body of the inlet and outlet valves strikes the valve seat ring at high speed. The result is on the one hand noise, and on the other the valve components are subject to relatively high wear. 
   In order to reduce the switching speed of the control valve shortly before the valve body strikes the valve seat ring, it is known to assign a hydraulically acting valve brake to the control valve piston. This valve brake is based on reducing a flow cross section for the hydraulic medium, so that a damping action ensues. A disadvantage, however, is that the braking action of the valve brake is very highly dependent on the viscosity of the hydraulic medium, which as a rule is hydraulic oil. The viscosity of the hydraulic medium is in turn highly temperature-dependent. As a result, the valve action of the valve brake and thus the impact speed of the valve body on the valve seat ring is highly temperature-dependent. 
   SUMMARY OF THE INVENTION 
   The electrohydraulic valve controller of the invention, conversely, offers the advantage that the impact speed of the valve body of the gas exchange valve on the valve seat can be reduced to a predeterminable constant value, virtually independently of any viscosity of the hydraulic medium. Because the valve brake includes a temperature compensation for the hydraulic medium, it is advantageously possible to compensate for fluctuating braking actions of the valve brake that are caused by temperature-dictated changes in viscosity. As a result, the impact speed of the valve body of the gas exchange valve can be set to a predeterminable value independently of any fluctuations in temperature. In particular, an automatic mechanical temperature compensation is possible as a result. 
   In a preferred feature of the invention, it is provided that the valve brake includes a first hydraulic circuit, forming a brake circuit, and a second hydraulic circuit, forming a compensation circuit; a hydraulic medium with essentially the same temperature is used in the brake circuit and in the compensation circuit. As a result, the temperature compensation is possible in an especially simple way, since if changes in temperature of the hydraulic medium occur in the brake circuit, the hydraulic medium in the compensation circuit undergoes the same change in temperature. Changes in viscosity caused by the temperature changes can thus be taken into account directly in the brake circuit, so that the braking action of the valve brake remains constant even if the temperatures fluctuate. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention will be explained in further detail below in conjunction with the associated drawings in which: 
     FIG.  1 . is a hydraulic circuit diagram of an electrohydraulic valve controller embodying the invention; and 
       FIG. 2  is a sectional view through a valve brake. 
   

   DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     FIG. 1  shows a circuit diagram of an electrohydraulic valve controller  10  for controlling a gas exchange valve  12 . The gas exchange valve  12  includes a valve body  14 , with which a valve seat embodied as a valve seat ring  16  is associated. The valve seat ring  16  is disposed in a cylinder head  18 , shown here only in suggested form, of an internal combustion engine. The structure and mode of operation of such gas exchange valves  12  are well known and therefore need not be addressed in detail in the context of the present description. 
   The valve controller  10  includes a hydraulic pumping device  20 , by means of which a hydraulic medium—hereinafter called hydraulic oil—can be pumped out of an oil sump  22  into a high-pressure reservoir  24 . The high-pressure reservoir  24  communicates with the oil sump  22  via a pressure limiting valve  26 , so that a defined oil pressure can be built up in the high-pressure reservoir  24 . 
   The high-pressure reservoir  24  moreover communicates via a check valve  28  with a bistable magnet valve  30  and a first pressure chamber  32  of a control valve  34 . The control valve  34  has a control valve piston  36 , which is guided tightly inside a cylinder  38 . Via an actuating means  40 , the control valve piston  36  is operatively connected to the valve body  14  of the gas exchange valve  12 . 
   The control valve piston  36  separates the first pressure chamber  32  of the control valve  34  from a second pressure chamber  42 . The second pressure chamber  42  communicates with the magnet valve  30  and, via a check valve  44 , with the high-pressure reservoir  24 . The second pressure chamber  42  also communicates via a hydraulic valve brake  46  with a second bistable magnet valve  48 . A conduit  50  also discharges into the cylinder  38  of the control valve  34  and communicates on its other end with the magnet valve  48 . The magnet valve  48  also communicates with a low-pressure reservoir  52 , which is in communication with the oil sump  22  via a check valve  54 . 
   The valve controller  10  shown in  FIG. 1  has the following function: 
   By means of the valve controller  10 , the gas exchange valve  12  can either be opened (not shown in  FIG. 1 ) or closed. Via the hydraulic pumping device  20 , a predeterminable pressure of the hydraulic oil is built up in the high-pressure reservoir  24 . By adjustment of the pressure limiting valve  26 , the level of this pressure can be determined. If an operating pressure that can be set by the check valve  28  is exceeded, the check valve  28  opens, and so hydraulic oil at this operating pressure is present in the pressure chamber  32  of the control valve  34 . For opening the gas exchange valve  12 , the magnet valves  30  and  48  are triggered in such a way that the magnet valve  30  is open, and the magnet valve  48  is closed. With the magnet valve  30  open, the operating pressure of the hydraulic oil also prevails in the pressure chamber  42 . Thus the same operating pressure prevails in both pressure chambers  32  and  42 . However, since the area of the control valve piston  36  acted upon by pressure in the pressure chamber  42  is greater than in the pressure chamber  32 , the control valve piston  36  is positively displaced in the direction of the pressure chamber  32 . As a result, the gas exchange valve  12  opens. The difference in surface area of the faces acted upon by pressure of the control valve piston  36  toward the pressure chamber  42  and toward the pressure chamber  32  is the result of the cross-sectional area of the actuating means  40  in the pressure chamber  32 . 
   Since the magnet valve  48  is closed, there is no communication with the low-pressure reservoir  52 . By the adjusting motion of the control valve piston  36 , the conduit  50  is opened toward the pressure chamber  42 , so that the valve brake  46  is idle and does not develop any action. 
   If the gas exchange valve  12  is to be closed, the magnet valves  30  and  48  are switched over; that is, the magnet valve  30  is closed, and the magnet valve  48  is open (as shown for these valves in FIG.  1 ). 
   With the magnet valve  30  closed, the operating pressure of the hydraulic oil prevails solely in the pressure chamber  32 . As a result, the control valve piston  36  is positively displaced in the direction of the pressure chamber  42 , until the valve body  14  of the gas exchange valve  12  strikes the valve seat ring  16 . During this adjusting motion of the control valve piston  36 , the conduit  50  is initially still open, so that the hydraulic oil located in the pressure chamber  42  is positively displaced into the low-pressure reservoir  52 . As soon as the upper control edge of the control valve piston  36  reaches the conduit  50 , this conduit is closed, so that the hydraulic oil from the pressure chamber  42  is positively displaced into the low-pressure reservoir  52  via the valve brake  46  and the magnet valve  48 . Thus by means of the valve brake  46 , just before the closing position of the gas exchange valve  12  is reached, a braking action ensues, so that the impact speed of the valve body  14  on the valve seat ring  16  is reduced. 
   The structure and mode of operation of the valve brake  46  will now be described in further detail in terms of the sectional view in FIG.  2 . 
   The valve brake  46  has a valve housing  60 , which forms an internal chamber  62 . The internal chamber  62  changes over from a larger-diameter portion  64  to a smaller-diameter portion  68  at an annular step  66 . A valve piston  70  is guided in the internal chamber  62 . The valve piston  70  has a piston or shoulder  72 , which has a smaller diameter than the portion  64  of the internal chamber  62 . As a result, between the shoulder  72  and the valve housing  60 , an annular gap  74  is formed, with a medium gap diameter dm that is the result of the difference between the diameter of the internal chamber  62  in the portion  64  and the diameter of the shoulder  72 . 
   An extension  76  that engages the portion  68  of the internal chamber  62  extends from the shoulder  72 . The extension  76  has a diameter that is equivalent to the diameter of the internal chamber  62  in the portion  68 . As a result, the extension  76  is guided sealingly in the portion  68 . A spring element  78  is braced on the annular step  66  and on the other end is supported on the shoulder  72 . 
   A first conduit  77  and a second conduit  79  discharge into the internal chamber  62  in the portion  68 . The conduit  77  is in communication with the pressure chamber  42  of the control valve  34 , and the conduit  79  is in communication with the magnet valve  48  (FIG.  1 ). In the region of the conduits  77  and  79 , the extension  76  has formed therein an annular groove  80 , and a bottom  82  of the annular groove  80  extends from a first control edge  84  to a second control edge  86 . The geometry of the bottom  82  is selected such that the conical tapering is only simplified, and the geometry of the bottom must be designed, as a function of the pressure difference p 1 −p 2 , the spring rate, and the viscosity behavior of the oil, such that the pressure drop in the brake circuit is always the same. 
   The conduit  77 , annular groove  80 , and conduit  79  form one brake circuit of the valve brake  46 . If the valve body  14  is to be braked and thus the control valve piston  36  is also to be braked, the hydraulic oil is present at the valve brake  46  via the conduit  77 . Depending on the position of the valve piston  70 , a throttle gap develops between the control edge  84  and the conduits  77  and  79 , respectively, by way of which gap the hydraulic oil reaches the annular groove  80 . The geometry of the annular groove  80  is designed such that the pressure in the pressure chamber  42  has no influence on the throttle gap and thus on the braking action (pressure compensation). 
   A further conduit  88  and a conduit  90  discharge into the internal chamber  62  in the region of the portion  64  of the internal chamber  62 . The conduit  90  discharges into the internal chamber  62  at an axial length from the conduit  88  that is greater than an axial length of the shoulder  72 . As a result, the conduits  88  and  90  are in fluidic communication with one another via the annular gap  74 . The conduit  88 , annular gap  74  and conduit  90  form a compensation circuit of the valve brake  46 . The conduit  90  communicates with the oil sump, so that in it a constant pressure p 2  is established. The compensation circuit is hydraulically disconnected from the brake circuit of the valve brake  46 . By suitable structural or other additional provisions that are not shown in detail, it is assured that the hydraulic oil in the compensation circuit has essentially the same temperature as the hydraulic oil in the brake circuit of the valve brake  46 . 
   The following relationships apply to the compensation circuit. Friction in the annular gap  74  causes a pressure loss Δ p, so that at the conduit  88 , the hydraulic oil of the compensation circuit is at a pressure p 1 ; the applicable equation is:
 
Δ p=p   1   −p   2 .
 
   A volumetric flow {dot over (V)} in the compensation circuit results in accordance with the following equation: 
           V   •     =       Δ   ⁢           ⁢     p   ·   s   ·   π   ·   d     ⁢           ⁢   m       12   ·   η   ·   l         ,       
 
in which s is the gap height, dm is the medium gap diameter, and l is the gap length of the annular gap  74 . The character η stands for the dynamic viscosity of the hydraulic oil in the compensation circuit. If all the factors that are dependent on the geometry of the annular gap  74  are combined into a geometry constant C, then the following equation applies: 
       C   =           s   ·   d     ⁢           ⁢     m   ·   π         12   ·   l       .         
 
   The result for the pressure loss is accordingly: 
         Δ   ⁢           ⁢   p     =           V   •     ·   η     C     .         
 
   Because of the pressures p 1  and p 2  and the force of the spring element  78 , the following force equilibrium ensues at the valve piston  70 :
 
 p   1   ·A 1 =p   2   ·A 2 +F, 
 
in which F is the spring force of the spring element  78 , and A 1  and A 2  are the areas, acted upon by pressure, of the shoulder  72  of the valve piston  70 . If this formula is solved for F, and if
 
 p   1   =Δp+p   2 
 
and if 
           Δ   ⁢           ⁢   p     =         V   •     ·   η     C       ,       
 
the result is 
         F   =           p   2     ·     (       A   ⁢           ⁢   1     -     A   ⁢           ⁢   2       )       +             V   •     ·   η     C     ·   A     ⁢           ⁢   2       =     R   ·   h         ,       
 
in which R is the spring rate and h is the spring height. For the spring height h, the result is accordingly: 
       h   =             p   2     ·     (       A   ⁢           ⁢   1     -     A   ⁢           ⁢   2       )       +             V   •     ·   η     C     ·   A     ⁢           ⁢   2       R     .         
 
   From this equation it becomes clear that the height h of the spring element  78  and thus the location of the valve piston  70  are directly dependent on the dynamic viscosity η of the hydraulic oil. If the dynamic viscosity η of the hydraulic oil changes, for instance because of a temperature change, then the location of the valve piston  70  changes automatically. The result is a compensation for a temperature-dependent change in viscosity of the hydraulic oil. 
   If the annular gap  74  of the spring element  78  and the annular groove  80  are suitably designed, it is accordingly possible to keep the impact speed of the valve body  14  on the valve seat ring  16  constant, independently of the instantaneous viscosity of the hydraulic oil. 
   The foregoing relates to preferred exemplary embodiments in 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.