Abstract:
The present invention provides a valve actuation system comprising a valve train for actuating a valve, the valve train including actuating elements and a valve lash, and a valve lash adjustment system for adjusting the valve lash, wherein the valve train and the valve lash adjustment system do not share any common actuating elements.

Description:
Priority is claimed under 35 U.S.C. §119(e) to U.S. Provisional Application No. 61/205,777 filed on Jan. 22, 2009, which is hereby incorporated by reference it its entirety. 
    
    
     TECHNICAL FIELD 
     The present invention relates generally to a valve lash adjustment system and a valve actuation system for a valve of an internal combustion engine. More specifically, the present invention relates to a valve lash adjustment system for a valve of a split-cycle engine. 
     BACKGROUND OF THE INVENTION 
     For purposes of clarity, the term “conventional engine” as used in the present application refers to an internal combustion engine wherein all four strokes of the well known Otto cycle (the intake, compression, expansion and exhaust strokes) are contained in each piston/cylinder combination of the engine. Each stroke requires one half revolution of the crankshaft (180 degrees crank angle (CA)), and two full revolutions of the crankshaft (720 degrees CA) are required to complete the entire Otto cycle in each cylinder of a conventional engine. 
     Also, for purposes of clarity, the following definition is offered for the term “split-cycle engine” as may be applied to engines disclosed in the prior art and as referred to in the present application. 
     A split-cycle engine comprises: 
     a crankshaft rotatable about a crankshaft axis; 
     a compression piston slidably received within a compression cylinder and operatively connected to the crankshaft such that the compression piston reciprocates through an intake stroke and a compression stroke during a single rotation of the crankshaft; 
     an expansion (power) piston slidably received within an expansion cylinder and operatively connected to the crankshaft such that the expansion piston reciprocates through an expansion stroke and an exhaust stroke during a single rotation of the crankshaft; and 
     a crossover passage interconnecting the compression and expansion cylinders, the crossover passage including a crossover compression (XovrC) valve and a crossover expansion (XovrE) valve defining a pressure chamber therebetween. 
     U.S. Pat. No. 6,543,225 granted Apr. 8, 2003 to Carmelo J. Scuderi (the Scuderi patent) and U.S. Pat. No. 6,952,923 granted Oct. 11, 2005 to David P. Branyon et al. (the Branyon patent) each contain an extensive discussion of split-cycle and similar type engines. In addition the Scuderi and Branyon patents disclose details of prior versions of engines of which the present invention comprises a further development. Both the Scuderi patent and the Branyon patent are incorporated herein by reference in their entirety. 
     Referring to  FIG. 1 , a prior art split-cycle engine of the type similar to those described in the Branyon and Scuderi patents is shown generally by numeral  10 . The split-cycle engine  10  replaces two adjacent cylinders of a conventional engine with a combination of one compression cylinder  12  and one expansion cylinder  14 . The four strokes of the Otto cycle are “split” over the two cylinders  12  and  14  such that the compression cylinder  12  contains the intake and compression strokes and the expansion cylinder  14  contains the expansion and exhaust strokes. The Otto cycle is therefore completed in these two cylinders  12 ,  14  once per crankshaft  16  revolution (360 degrees CA). 
     During the intake stroke, intake air is drawn into the compression cylinder  12  through an inwardly opening (opening inward into the cylinder) poppet intake valve  18 . During the compression stroke, compression piston  20  pressurizes the air charge and drives the air charge through the crossover passage  22 , which acts as the intake passage for the expansion cylinder  14 . 
     Due to very high volumetric compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the compression cylinder  12 , an outwardly opening (opening outward away from the cylinder) poppet crossover compression (XovrC) valve  24  at the crossover passage inlet is used to control flow from the compression cylinder  12  into the crossover passage  22 . Due to very high volumetric compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the expansion cylinder  14 , an outwardly opening poppet crossover expansion (XovrE) valve  26  at the outlet of the crossover passage  22  controls flow from the crossover passage  22  into the expansion cylinder  14 . The actuation rates and phasing of the XovrC and XovrE valves  24 ,  26  are timed to maintain pressure in the crossover passage  22  at a high minimum pressure (typically 20 bar or higher) during all four strokes of the Otto cycle. 
     A fuel injector  28  injects fuel into the pressurized air at the exit end of the crossover passage  22  in correspondence with the XovrE valve  26  opening. The fuel-air charge fully enters the expansion cylinder  14  shortly after expansion piston  30  reaches its top dead center position. As piston  30  begins its descent from its top dead center position, and while the XovrE valve  26  is still open, spark plug  32  is fired to initiate combustion (typically between 10 to 20 degrees CA after top dead center of the expansion piston  30 ). The XovrE valve  26  is then closed before the resulting combustion event can enter the crossover passage  22 . The combustion event drives the expansion piston  30  downward in a power stroke. Exhaust gases are pumped out of the expansion cylinder  14  through inwardly opening poppet exhaust valve  34  during the exhaust stroke. 
     With the split-cycle engine concept, the geometric engine parameters (i.e., bore, stroke, connecting rod length, compression ratio, etc.) of the compression and expansion cylinders are generally independent from one another. For example, the crank throws  36 ,  38  for the compression cylinder  12  and expansion cylinder  14  respectively may have different radii and may be phased apart from one another with top dead center (TDC) of the expansion piston  30  occurring prior to TDC of the compression piston  20 . This independence enables the split-cycle engine to potentially achieve higher efficiency levels and greater torques than typical four stroke engines. 
     The actuation mechanisms (not shown) for crossover valves  24 ,  26  may be cam driven or camless. In general, a cam driven mechanism includes a camshaft mechanically linked to the crankshaft. A cam is mounted to the camshaft, and has a contoured surface that controls the valve lift profile of the valve opening event [i.e., the event that occurs during a valve actuation]. A cam driven actuation mechanism is efficient, fast and may be part of a variable valve actuation system, but generally has limited flexibility. 
     For purposes herein a valve opening event is defined as the valve lift from its initial opening off of its valve seat to its closing back onto its valve seat versus rotation of the crankshaft during which the valve lift occurs. Also for purposes herein the valve opening event rate [i.e., the valve actuation rate] is the duration in time required for the valve opening event to occur within a given engine cycle. It is important to note that a valve opening event is generally only a fraction of the total duration of an engine operating cycle, e.g., 720 CA degrees for a conventional engine cycle and 360 CA degrees for a split-cycle engine. 
     Also in general, camless actuation systems are known, and include systems that have one or more combinations of mechanical, hydraulic, pneumatic, and/or electrical components or the like. Camless systems allow for greater flexibility during operation, including, but not limited to, the ability to change the valve lift height and duration and/or deactivate the valve at selective times. 
     Referring to  FIG. 2 , an exemplary prior art valve lift profile  40  for a crossover valve in a split-cycle engine is shown. Valve lift profile  40  can potentially be applied to either or both of crossover valves  24 ,  26  in  FIG. 1 . Valves  24  and  26  will be referred to below as having the same valve lift profile  40  merely for purposes of discussion. 
     Regardless of whether valves  24  and  26  are cam driven or actuated with a camless system, the valve lift profile  40  needs to be controlled to avoid damaging impacts when the valves  24 ,  26  are approaching their closed positions against their valve seats. Accordingly, a portion of the profile  40 —referred to herein as the “landing” ramp  42 —may be controlled to rapidly decelerate the velocity of the valves  24 ,  26  as they approach their valve seats. The valve lift at the start of maximum deceleration (on the descending side of the profile  40 ) is defined herein as the landing ramp height  44 . The landing ramp duration  46  is defined herein as the duration of time from the start of the maximum deceleration of the moving valve to the point of landing on the valve seat. The velocity of the valve  24  or  26  when the valve contacts the valve seat is referred to herein as the seating velocity. For purposes herein, the “takeoff” ramp  45  is not as critical as the landing ramp  42 , and can be set to any value that adequately achieves the maximum lift  48 . 
     In cam-driven actuation systems, the landing ramp is generated by the profile of the cam. Accordingly, the landing ramp&#39;s duration in time is proportional to the engine speed, while its duration relative to crankshaft rotation (i.e., degrees CA) is generally fixed. In camless actuation systems, in general, the landing ramp is actively controlled by a valve seating control device or system. 
     For split-cycle engines which ignite their charge after the expansion piston reaches its top dead center position (such as in the Scuderi and Branyon patents), the dynamic actuation of the crossover valves  24 ,  26  is very demanding. This is because the crossover valves  24  and  26  of engine  10  must achieve sufficient lift to fully transfer the fuel-air charge in a very short period of crankshaft rotation (generally in a range of about 30 to 60 degrees CA) relative to that of a conventional engine, which normally actuates the valves for a period of at least 180 degrees CA. This means that the crossover valves  24 ,  26  must actuate about four to six times faster than the valves of a conventional engine. 
     As a consequence of the faster actuation requirements, the XovrC and XovrE valves  24 ,  26  of the split-cycle engine  10  have a severely restricted maximum lift ( 48  in  FIG. 2 ) compared to that of valves in a conventional engine. Typically the maximum lift  48  of these crossover valves  24 ,  26  are in the order of 2 to 3 millimeters, as compared to about 10-12 mm for valves in a conventional engine. Consequently, both the height  44  and duration  46  of the landing ramp  42  for the XovrC and XovrE valves  24 ,  26 , need to be minimized to account for the shortened maximum lift and faster actuation rates. 
     Problematically, the heights  44  of the landing ramps  42  of crossover valves  24  and  26  are so restricted that unavoidable variations in parameters that control ramp height, which are normally less significant in their effect on the larger lift profiles of conventional engines, now become critical. These parameter variations may include, but are not limited to:
         1) dimensional changes due to thermal expansion of the metal valve stem and other metallic components in the valve&#39;s actuation mechanism as engine operational temperatures vary;   2) the normal wear of the valve and valve seat during the operational life of the valve;   3) manufacturing and assembly tolerances; and   4) variations in the compressibility (and resulting deflection) of hydraulic fluids (e.g. oil) in any components of the valvetrain (mainly caused by aeration).       

     Referring to  FIG. 3 , an exemplary embodiment of a conventional cam-driven valve train  50  for a conventional engine is illustrated. For purposes herein, a valve train of an internal combustion engine is defined as a system of valve train elements, which is used to control the actuation of the valves. The valve train elements generally comprise a combination of actuating elements and their associated support elements. Also for purposes herein, the primary motion of any valve train element is defined as that motion which the element would substantially experience when the elements of the valve train are idealized to have an infinite stiffness. The actuating elements (e.g., cams, tappets, springs, rocker arms, valves and the like) are used to directly impart the primary actuation motion to the valves (i.e., to actuate the valves) of the engine during each valve opening event of the valves. Accordingly, the primary motion of the individual actuating elements in a valve train must operate at the substantially same actuation rates as the valve opening events of the valves that the actuating elements actuate. The support elements (e.g., shafts, pedestals or the like) are used to securely mount and guide the actuating elements to the engine and generally have no primary motion, although they affect the overall stiffness of the valve train system. However, the primary motion, if any, of the support elements in a valve train operate at slower rates than the valve opening events of the valves. 
     It should be noted that support elements may be subject to some high frequency vibration primarily caused by the high frequency movements of the actuating elements of a valve train, which apply forces to the support elements during operation. The high frequency vibrations are a consequence of the actuating and support elements of the valve train having a finite stiffness, and are not part of the primary motion. However, the displacement induced by this vibration alone will have a magnitude that is substantially less than the magnitude of the primary motion of the actuating elements in the valve train, typically by an order of magnitude or less. 
     Valve train  50  actuates an inwardly opening poppet valve  52  having a valve head  54  and a valve stem  56 . Located at the distal end of the valve stem  56  is the valve tip  58 , which abuts against a tappet  60 . Spring  62  holds the valve head  54  securely against a valve seat  64  when the valve  52  is in its closed position. Cam  66  rotates to act against the tappet  60  in order to depress spring  62  and lift the valve head  54  off of its valve seat  64 . In this exemplary embodiment, valve  52 , spring  62 , tappet  60  and cam  66  are actuating elements. Though no associated support elements are illustrated, one skilled in the art would recognize that they would be required. Cam  66  includes a cylindrical portion, generally referred to as the base circle  68 , which does not impart any linear motion to the valve  52 . Cam  66  also includes a lift (or eccentric) portion  70  that imparts the linear motion to the valve  52 . The contour of the cam&#39;s eccentric portion  70  controls the lift profile of valve  52 . The effects of the aforementioned dimensional changes due to thermal expansion are compensated for by including a preset clearance space (or clearance)  72 . 
     For purposes herein, the terms “valve lash” or “lash”: are defined as the total clearance existing within a valve train when the valve is fully seated. The valve lash is equal to the total contribution of all the individual clearances between all individual valve train elements (i.e., actuating elements and support elements) of a valve train 
     In this particular embodiment, the clearance  72  is the distance between the base circle  68  of cam  66  and the tappet  60 . Also note that, in this particular embodiment, the clearance  72  is substantially equal to the valve lash of the valve train, i.e., the total contribution of all the clearances that exist between the valve&#39;s distal tip  58 , when the valve  52  is fully seated on the valve seat  64 , and the cam  66 . 
     To compensate for the thermal effects on the inwardly opening valve  52 , the clearance  72  is set at its maximum tolerance when the engine is cold. When the engine heats up, the valve&#39;s stem  56  will expand in length and reduce the clearance  72 , but will not abut against the cam&#39;s base circle  68  (i.e., will not reduce the clearance  72  to zero). Accordingly, as the clearance  72  is reduced, valve  52  is extended further into the cylinder (not shown) when the valve  52  is open. Note however that, even as the clearance  72  is reduced, valve  52  remains seated against its valve seat when the valve  52  is closed. 
     However, as mentioned above, crossover valves, such as valves  24 ,  26  in split-cycle engine  10 , have lift profiles that include much smaller landing ramp heights compared to that of a conventional engine. This would be true whether the valves were inwardly opening or outwardly opening, so long as the duration of valve actuation [i.e., the valve opening event] was short relative to that of a valve on a conventional engine, for example, a valve with a duration of actuation of approximately 3 ms and 180 degrees of crank angle, or less. In the case of such fast actuating, cam driven, inwardly opening valves, the valve&#39;s distal tip must engage the cam&#39;s landing ramps in order to have a controlled landing and safe seating velocity, and any fixed valve lash for such inwardly opening crossover valves must necessarily be set proportionally small. Problematically, variations in a set valve lash due to thermal expansion effects may actually be greater than the ramp height required for such valves. This means that if the valve lash is set large enough to account for thermal expansion, the tips of these inwardly opening crossover valves could miss the landing ramp altogether, which would cause the valves to repeatedly crash against their valve seats and prematurely damage the valves. Additionally, if the valve lash is set small enough to guarantee engagement with the landing ramp at all operating temperatures, the tips of the valves could expand enough to abut against the base circle of the cam, which would force the inwardly opening crossover valves open even when the valves should be in their closed position. 
     Moreover, the large lash setting would generate a shorter valve lift duration and the small lash setting would generate a lengthened valve lift duration. In either case, the range of variation of the valve opening event can be larger than desirable. It is desirable to contain the range of the valve opening event to a manageable level. 
     Referring to  FIG. 4 , an exemplary embodiment of a conventional engine cam driven valve train  73  having an automatically adjustable valve lash is illustrated. The valve train  73  actuates inwardly opening poppet valve  74 . The valve train  73  includes cam  76 , pivoting lever arm  78  and spring  80  as valve train actuating elements which actuate valve  74  during each cycle. The effects of thermal expansion and other parameters mentioned above are addressed by adding a lash adjuster assembly. For the lash adjuster assembly, an active lash control device, such as a hydraulic lash adjuster (HLA)  82  has been used. The hydraulic lash adjuster (HLA)  82  also functions as a support element associated with lever arm  78 . As is known in the art, as valve lash in the valve train varies, HLA  82  hydraulically adjusts the position of lever arm  78  to compensate and bring the valve lash to zero (in this particular embodiment, the valve lash would be any clearance between the cam  76  and the lever arm  78 , as well as any clearance between the lever arm  78  and the distal tip of the stem of valve  74 ). 
     Because lever arm  78  is one of the valve train  73  actuating elements (i.e., is an element that directly actuates the inwardly opening valve  74  during each cycle and is used to directly impart the primary actuation motion to the valve  74 ), there is an unavoidable tradeoff between the lever arm&#39;s minimum mass required for adequate stiffness (ratio of force applied to a point on the lever arm to the deflection of that point caused by that force) and the maximum mass allowable for high speed operation. That is, if the mass of lever arm  78  is too small, it will not be able to actuate valve  74  without undue bending and/or deformation. Additionally, if the mass of lever arm  78  is too large, it will be too heavy to actuate valve  74  at its maximum operating speed. For any particular valve train actuating element, if the minimum mass required for adequate stiffness exceeds the maximum mass allowable for maximum operating speed, the element cannot be used in the valve train. Generally, in a conventional engine, the requirements for stiffness and speed are not so demanding as to preclude the use of lever arm  78  in valve train  73 . 
     However, as mentioned above, crossover valves  24 ,  26  must actuate about four to six times faster than the valves of a conventional engine, which means the actuating elements of the valve train system must operate at extremely high and rapidly changing acceleration levels relative to that of a conventional engine. These operating conditions would severely restrict the maximum mass of lever arm  78  in valve train  73 . 
     Additionally, crossover valves  24 ,  26  must open against very high pressures in the crossover passage  22  compared to that of a conventional engine (e.g., 20 bar or higher), which exacerbates the stiffness requirements on the valve train system. Also, bending is a problem on elements such as lever arm  78  because the actuation force in one direction is concentrated in the median section of the element (i.e., where cam  76  engages lever arm  78 ) and all opposing reactionary forces are concentrated at the end sections of the lever arm (i.e., where HLA  82  and the tip of valve  74  engage opposing ends of lever arm  78 ). Moreover, this bending problem would increase proportionally as the length of the lever arm  78  increases. Accordingly, if the engine illustrated in prior art  FIG. 4  were subjected to the higher pressures and severe actuation rates encountered in split-cycle engine  10 , the stiffness and mass of lever arm  78  in valve train  73  would have to be substantially increased, therefore restricting the overall actuation rate of valve train  73 . 
     Generally too, prior art HLAs (such as HLA  82 ), because of the compressibility of oil contained therein, are normally one of the main contributing factors in reducing valve train stiffness which, in turn, limits the maximum engine operating speed at which the valve train can safely operate. Therefore, a prior art HLA  82  connected to a lever arm  78 , as shown in valve train  73 , cannot be implemented with the split cycle engine  10 , in which the valves need to actuate much more rapidly, and the HLA  82  must be much stiffer than those in a conventional engine. 
     There is a need therefore, for a valve lash adjustment system for cam driven valves of a split-cycle engine, which can both (a) handle the high speed and stiffness requirements necessary to safely actuate the valves; and (b) automatically compensate for such unavoidable factors as thermal expansion of actuation components, valve wear, and manufacturing tolerances that cause variations in the lash. 
     SUMMARY OF THE INVENTION 
     A valve actuation system ( 150 ) comprising a valve train ( 152 ) for actuating a valve ( 132 / 134 ), the valve train ( 152 ) including actuating elements ( 161 ,  162 ,  132 / 134 ) and a valve lash ( 178 ,  180 ); and a valve lash adjustment system ( 160 ) for adjusting the valve lash ( 178 ,  180 ), wherein said valve train ( 152 ) and said valve lash adjustment system ( 160 ) do not share any common actuating elements. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic cross-sectional view of a prior art split-cycle engine related to the engine of the invention; 
         FIG. 2  shows an exemplary prior art valve lift profile for a cross-over valve in a split-cycle engine; 
         FIG. 3  shows a prior art cam-driven valve train of a conventional engine; 
         FIG. 4  is a schematic cross-sectional view of a prior art hydraulic valve lash adjustment system, which uses a finger lever pivot element 
         FIG. 5  shows an exemplary embodiment of the valve lash adjustment system of the invention mounted on a split-cycle engine; 
         FIGS. 6 ,  7  and  8  show a side view, perspective view and exploded view, respectively, of an exemplary embodiment of the valve lash adjustment system and valve train of the invention; 
         FIG. 9  shows an exploded view of some of the key components of the valve lash adjustment system; 
         FIG. 10  is a perspective view of the rocker of the valve train only, and the rocker shaft of both the valve lash adjustment system and valve train; 
         FIG. 11  is a top view of the rocker shaft and rocker shaft lever of the valve lash adjustment system; 
         FIGS. 12 and 13  show the motion of the rocker arm of the valve lash adjustment system; and 
         FIG. 14  is an enlarged view of center section  14 - 14  of  FIG. 13 . 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Referring to  FIG. 5 , numeral  100  generally indicates a diagrammatic representation of an exemplary embodiment of a split-cycle engine according to the present invention. Engine  100  includes a crankshaft  102  rotatable about a crankshaft axis  104  in a clockwise direction as shown in the drawing. The crankshaft  102  includes adjacent angularly displaced leading and following crank throws  106 ,  108 , connected to connecting rods  110 ,  112 , respectively. 
     Engine  100  further includes a cylinder block  114  defining a pair of adjacent cylinders, in particular a compression cylinder  116  and an expansion cylinder  118  closed by a cylinder head  120  at one end of the cylinders opposite the crankshaft  102 . A compression piston  122  is received in compression cylinder  116  and is connected to the connecting rod  112  for reciprocation of the piston  122  between top dead center (TDC) and bottom dead center (BDC) positions. An expansion piston  124  is received in expansion cylinder  118  and is connected to the connecting rod  110  for similar TDC/BDC reciprocation. The diameters of the cylinders  116 ,  118  and pistons  122 ,  124  and the strokes of the pistons  122 ,  124  and their displacements need not be the same. 
     Cylinder head  120  provides the means for gas flow into, out of and between the cylinders  116  and  118 . The cylinder head  120  includes an intake port  126  through which intake air is drawn into the compression cylinder  116  through an inwardly opening poppet intake valve  128  during the intake stroke. During the compression stroke, compression piston  122  pressurizes the air charge and drives the air though a crossover (Xovr) passage  130 , which acts as the intake passage for the expansion cylinder  118 . 
     Due to very high compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the compression cylinder  116 , an outwardly opening poppet crossover compression (XovrC) valve  132  at the crossover passage inlet is used to control flow from the compression cylinder  116  to the crossover passage  130 . Due to very high compression ratios (e.g., 20 to 1, 30 to 1, 40 to 1, or greater) within the expansion cylinder  118 , an outwardly opening poppet crossover expansion (XovrE) valve  134  at the outlet of the crossover passage  130  controls flow from the crossover passage  130  into the expansion cylinder  118 . Crossover compression valve  132 , crossover expansion valve  134  and crossover passage  130  define a pressure chamber  136  in which pressurized gas (typically 20 bar or greater) is stored between closing of the crossover expansion (XovrE) valve  134  during the expansion stroke of the expansion piston  124  on one cycle (crank rotation) of the engine  100  and opening of the crossover compression (XovrC) valve  132  during the compression stroke of the compression piston  122  on the following cycle (crank rotation) of the engine. 
     A fuel injector  138  injects fuel into the pressurized air at the exit end of the crossover passage  130  in correspondence with the XovrE valve  134  opening. The fuel-air charge enters the expansion cylinder  118  shortly after expansion piston  124  reaches its top dead center position. As piston  124  begins its descent from its top dead center position, and while the XovrE valve  134  is still open, spark plug  140  is fired to initiate combustion (typically between 10 to 20 degrees CA after top dead center of the expansion piston  124 ). The XovrE valve  134  is then closed before the resulting combustion event can enter the crossover passage  130 . The combustion event drives the expansion piston  124  downward in a power stroke. Exhaust gases are pumped out of the expansion cylinder  118  through inwardly opening poppet exhaust valve  142  during the exhaust stroke. 
     The actuation mechanisms (not shown) for inlet valve  128  and exhaust valve  142  may be any suitable cam driven or camless system. Crossover compression and crossover expansion valves  132 ,  134  may also be actuated in any suitable manner. However, in accordance with the invention, preferably both crossover valves  132  and  134 , are actuated by a cam-driven actuation system  150 . Actuation system  150  comprises a valve train  152  that includes required actuating elements that are used to directly impart the primary actuation motion to the valves  132 ,  134 , and a separate valve lash adjustment system  160  mounted remotely from the valve train  152 . More specifically, the valve lash adjustment system  160  includes no actuating elements that are shared with the valve train  152 , and no element of the lash adjustment system  160  is used to directly impart the primary actuation motion of the valves  132  and  134 . 
     Referring to  FIGS. 6 ,  7  and  8 , a side view, perspective view and exploded view respectively of an exemplary embodiment of the cam driven actuation system  150  for crossover valves  132  and  134  are shown. 
     Referring to  FIGS. 6 and 7 , the valve train  152  for each crossover valve  132 ,  134  includes the cam  161 , rocker  162  and crossover valves  132 / 134  as actuating elements. As shown in  FIG. 8 , each of the valves  132 / 134  includes a valve head  164  and a valve stem  166  extending vertically from the valve head. A collet retainer  168  is disposed at the distal tip  169  of the stem  166  and securedly fixed thereto with a collet  170  and clip  172 . 
     Referring to  FIG. 8 , the rocker  162  includes a forked rocker pad  174  at one end, which straddles valve stem  166  and engages the underside of collet retainer  168 . Additionally, rocker  162  also includes a solid rocker pad  176  at an opposing end, which slidingly contacts cam  161  of the valve train  152 . Additionally, rocker  162  includes a rocker shaft bore  177  extending therethrough (see more detailed discussion below). 
     The forked rocker pad  174  of the rocker  162  contacts the collet retainer  168  of the outwardly opening poppet valves  132 / 134  such that a downward direction of the rocker pad  176  (direction A in  FIGS. 6 ,  12  and  13 ) caused by the actuation of the cam  161  translates into an upward movement of the rocker pad  174  (direction B in  FIGS. 6 ,  12  and  13 ), which opens the valves  132 / 134 . A gas spring (not shown) acts on the valves  132 / 134  to keep the valves  132 / 134  closed when not driven by the rocker  162 . 
     As shown in  FIG. 6 , valve lash in valve train  152  includes, but is not limited to, any clearances between the rocker  162  and the cam  161  and between the rocker  162  and the collett retainer  168  of the valves  132 ,  134 . Specifically, clearance  178  is the clearance between collet retainer  168  and rocker pad  174 . Additionally, clearance  180  is the clearance between cam  161  and rocker pad  176 . In this embodiment, element clearances  178  and  180  substantially comprise the valve lash of the valve train  152 . As will be explained herein below, valve lash adjustment system  160  adjusts the clearances  178  and  180  to a substantially zero clearance, and, therefore, adjusts the valve lash of valve train  152  to substantially zero. 
     In the present invention, the elements of the valve lash adjustment system  160  are mounted remotely relative to the valve train  152  in order to increase stiffness of the valve lash adjustment system, as explained further below. More specifically, no element of the valve lash adjustment system  160  is also an actuating element of the valve train  152 , and no element of the valve lash adjustment system  160  is configured to directly impart primary actuation motion to the valves  132  and  134 . As a result, the primary motion, if any, of the individual elements of the valve lash adjustment system  160  operate at slower rates than the actuation rates of valves  132  and  134 . As shown in  FIGS. 8 and 9 , the valve lash adjustment system  160  includes rocker shaft assembly  200 , which rotatably supports the rocker  162  of valve train  152 , a rocker shaft lever  300 , a pedestal assembly  400 , which rotatably contains the rocker shaft assembly  200 , and a lash adjuster assembly  600 . In this exemplary embodiment, a hydraulic lash adjuster (HLA) assembly is used as the lash adjuster assembly  600 . It should be noted that the HLA assembly is specific to this exemplary embodiment. One skilled in the art would recognize that other lash adjustment assemblies may used, e.g., pneumatic, mechanical or electrical lash adjust assemblies, or the like. 
     It is important to note that both the rocker shaft assembly  200  and the pedestal assembly  400 , of the valve lash adjustment system  160 , are also support elements of the valve train  152 . That is, the pedestal assembly  400  and the rocker shaft assembly  200  both provide support for the rocker  162  and affect the overall stiffness of the valve train  152 . However, the pedestal assembly  400  and the rocker shaft assembly  200  are not required to cycle at the same actuation rates or relative amplitudes as the actuating elements of valve train  152 . 
     As best seen in  FIG. 10 , the valve lash adjustment system  160  engages the valve train  152  only at the rocker  162 . That is, rocker  162  pivotally rotates on a relatively stationary rocker shaft assembly  200 . Note that rocker  162  is an element of the valve train  152  and is not an element of the valve lash adjustment system  160 , whereas rocker shaft assembly  200  is both an element of the valve lash adjustment system  160  and a support element of the valve train  152 . Accordingly, the rocker shaft assembly  200  does not directly impart primary actuation motion to valves  132  and  134  as an actuating element would, but rather acts as a relatively stationary shaft upon which rocker  152  pivots to actuate valves  132  and  134 . 
     As best seen in  FIGS. 8 and 9 , the pedestal assembly  400  includes pedestal  402  that is rigidly secured to the engine block (not shown), for example with bolts  404 , or other similar fasteners. The pedestal assembly  400  also includes a pedestal shim  406  having a predetermined thickness for accurately positioning the pedestal  402  relative to the valve train  152  in the vertical direction (direction of travel of valves  132 ,  134 ). Solid dowel  408  and hollow dowel  410  are utilized to accurately align the pedestal  402  relative to the valve train  152  in the horizontal direction. 
     Pedestal  402  has machined therein a front wall  412  and rear wall  414  defining a slot  416  therebetween. The pedestal slot  416  is sized to receive therein the rocker  162 . The front wall  412  and rear wall  414  include a front bore  418  and a rear bore  420  formed therein respectively. Front and rear bores  418 ,  420  are concentric around a fixed axis  422 , best shown in  FIG. 9 . Front and rear bores  418 ,  420  are sized to receive the rocker shaft assembly  200 , as described in detail below. 
     The rocker shaft assembly  200  includes a rocker shaft  202  and an eccentric rocker shaft cap  204  that is fixedly secured to the rocker shaft  202  via pins  207  and bolt  320 . The rocker shaft  202  includes a pedestal bearing portion  206  sized to be slip fit into front bore  418  such that the pedestal bearing portion  206  is concentric to the fixed axis  422 . The rocker shaft  202  also includes a rocker bearing portion  208  which is sized to be received in the rocker bore  177  such that the rocker  162  rotates and pivots on the rocker bearing portion  208 . When the rocker  162  is mounted onto the rocker bearing portion  208  with the rocker  162  inserted into the slot  416  formed in the pedestal  402  and the pedestal bearing portion  206  of the rocker shaft  202  is captured by the front bore  418 , the rocker  162  rotates about rocker bearing portion  208  within the slot  416 . As shown in  FIG. 9 , rocker bearing portion  208  is eccentric to the pedestal bearing portion  206  such that a center line of the rocker bearing portion  208  (the movable rocker axis  210 ) is offset from the fixed axis  422  by approximately 2 mm. Because the rocker  162  rotates on the rocker bearing portion  208 , the rocker  162  rotates about this movable rocker axis  210  as it actuates the valves  132 ,  134 . 
     Eccentric cap  204  includes an outer bearing surface  212  sized to slip fit into the rear bore  420  of the rear wall  414  of the pedestal  402  such that the outer bearing surface  212  is concentric with the fixed axis  422 . Eccentric cap  204  additionally includes an eccentric inner bearing surface  214  that receives and captures the rocker bearing portion  208 . The inner bearing surface  214  is concentric with the movable rocker axis  210 . 
     Because the rocker bearing portion  208  is eccentric to the pedestal bearing portion  206  and the outer bearing surface  212 , the rotation of the pedestal bearing portion  206  about the fixed axis  422  causes the rocker bearing portion  208  to move eccentrically with respect to the pedestal bearing portion  206  and the outer bearing surface  212 . That is, the rotation of the pedestal bearing portion  206  about the fixed axis  422  (best seen in  FIG. 14 ) causes the center of the rocker bearing portion  208  (the movable rocker axis  210 ) to move arcuately about the fixed axis  422 , as described in more detail below with respect to  FIGS. 12 ,  13  and  14 . Since the rocker  162  rotates on the rocker bearing portion  208 , this movement of the center  210  of the rocker bearing portion  208  adjusts the position of the rocker pad  176  relative to the cam  161 , and the position of the rocker pad  174  relative to the collet retainer  168 , thereby controlling the clearances  180 ,  178  and, therefore, the valve lash of valve train  152 . 
     The rotational angle of the rocker shaft assembly  200  is controlled by the rocker shaft lever  300 , to which it is rigidly joined by screw  320  or other similar fastener. As best shown in  FIG. 11 , the screw  320  is aligned with the movable rocker axis  210 . As shown in  FIGS. 8 and 9 , the rocker shaft lever  300  is coupled to the hydraulic lash adjuster (HLA) assembly  600  so that the rotational position of the rocker shaft lever  300  is controlled by the vertical deflection of the hydraulic lash adjuster (HLA) assembly  600 . The HLA assembly  600  includes a connecting cap  610  that is disposed on an upper end of a hydraulic lash adjuster  620  (HLA  620 ). The connecting cap  610  includes a pin  608  extending vertically from a base  606 . The base  606  further includes an upper surface  607  and a lower generally spherically-shaped socket  609 . The pin  608  is contained in a clearance slot  310  of the rocker shaft lever  300 . The lower socket  609  fits onto a generally spherically-tipped plunger  630  such that the cap  610  is free to rotate on the plunger  630 . The upper surface  607  of cap  610  abuts flush against a lower surface of rocker shaft lever  300  such that the cap  610  is captured between the lever  300  and HLA plunger  630 . Note that pin  608  is primarily used for ease of assembly and is not required to capture cap  610 . Clip  611  is optionally fitted to further assist assembly. Pressurized hydraulic fluid (not shown) is fed into HLA  620  to extend plunger  630  which raises connecting cap  610 , thereby rotating rocker shaft lever  300 . End  640  of the hydraulic lash adjuster (HLA) assembly  600  is mounted to the cylinder head (not shown) as is well known. For the hydraulic lash adjuster  620 , a Schaeffler F-56318-37 finger lever pivot element, or any other similar pivot element, can be used. As mentioned above, a hydraulic lash adjuster (HLA) assembly is used as the lash adjuster assembly  600  in this exemplary embodiment. It should be noted that the HLA assembly is specific to this exemplary embodiment. One skilled in the art would recognize that other lash adjustment assemblies may used, e.g., pneumatic, mechanical or electrical lash adjust assemblies, or the like. 
     Since the rocker  162  is part of the valve train  152 , it must be made very stiff. Also, because the rocker  162  is subjected to the high frequency actuation motion of the drive train, its mass must be minimized. Accordingly, the rocker  162  is machined from steel or stiffer materials and includes reinforcing ribs, as shown in  FIG. 10 . The configuration of the rocker  162  can be determined by performing well-known finite element analysis calculations. 
     As shown best in  FIG. 9 , the rocker shaft assembly  200  includes a male connecting portion  216  attached to the pedestal bearing portion  206 , which fits into a female connecting portion formed in the rocker shaft lever  300  so that the rocker shaft lever  300  and the rocker shaft assembly  200  rotate together about fixed axis  422 . Therefore, translational movement of the plunger  630  along axis  612  causes rotation of the rocker shaft assembly  200 . This rotation of the rocker shaft assembly  200  causes displacement of the rocker  162 , which is coupled to the rocker bearing portion  208  of the rocker shaft assembly  200 , as presented above. 
     The shape and orientation of the male connecting portion  216  of the rocker shaft assembly  200  and the corresponding shape and orientation of the female connecting portion of the rocker shaft lever  300  determine the orientation of the rocker shaft lever  300  relative to the rocker shaft assembly  200 . 
     As shown in  FIGS. 12 ,  13  and  14 , pressurized hydraulic fluid feeding into the HLA  620  causes the plunger  630  to extend outwardly toward a fully extended position from a fully retracted position relative to HLA  620 . This results in the rotation of the rocker shaft lever  300 , which causes an arcuate movement (as indicated by directional arrow  220  in  FIGS. 13 and 14 ) of the movable rocker axis  210  of the rocker bearing portion  208  about the fixed axis  422 . As can be best seen in  FIG. 14 , this arcuate movement  220  has both a vertical and horizontal component of direction. This results in a displacement of the rocker pad  176  of the rocker  162  towards the cam  161 , and displacement of the rocker pad  174  towards collet retainer  168 , thereby reducing the clearances  180  and  178  to substantially zero, as shown in  FIG. 13 . Accordingly, the valve lash, of which clearances  180  and  178  substantially comprise, is also reduced to substantially zero. 
     The embodiments described above describe a valve lash adjustment system  160  which reduces the lash to substantially zero, wherein there is contact between the cam  161  and the pad  176  of the rocker  162 , which causes frictional drag. This contact between the cam  161  and the pad  176  will drain energy from the engine. Therefore, it may be desirable to include a friction reduction mechanism (not shown) to either reduce frictional drag or limit the lash to some non-zero minimum value in order to prevent contact between the cam  161  and the pad  176  of the rocker  162 . 
     One such mechanism could be a non-rotating disc mounted to the camshaft by a bearing which holds the rocker pad  176  off of the base circle of the cam  161 . Alternatively a fixed stop or rest for the rocker  162  could be rigidly mounted to the cylinder head  120  to separate the rocker pad  176  from the base circle of the cam  161 . In the case of both the non-rotating disc and the fixed stop, it may be desirable that they have a coefficient of expansion approximately equal to the coefficient of expansion of the cam  161  to take into account the effects of thermal expansion. Alternatively, a roller could be added to the rocker pad  176  to reduce frictional drag between rocker  162  and cam  161 . 
     For purposes herein, the following definitions will be referred to and applied:
         1) stiffness (K 600 ) of the HLA assembly  600 : the ratio of the force (F 600 ) applied to the HLA plunger  630  (by the rocker shaft lever  300 ) to the deflection (D 600 ) of the plunger  630  (in the direction of the applied force) directly caused by the application of that force; and   2) stiffness (K 200 ) of the rocker shaft assembly  200 : the ratio of the force (F 200 ) applied to the rocker shaft assembly  200  by the rocker  162  to the deflection (D 200 ) of the rocker shaft assembly  200  (in the direction of the applied force) directly caused by the application of that force.
 
The stiffness of the rocker shaft assembly  200 , i.e., K 200 , can be subdivided into the following two main components:
   (A) the bending component (K 200 B), caused primarily by the deflection (D 200 B) resulting from the deformation of the various components of the rocker shaft assembly  200 , but primarily due to the bending of rocker bearing portion  208 ; and   (B) the rotating component (K 200 R), caused primarily by the deflection (D 200 R) resulting from the rotation of rocker shaft assembly  200  produced by the deflection of HLA assembly  600 .
 
Additionally, the approximate relationship between K 200 R and K 200 B is as follows: 1/K 200 =1/K 200 R+1/K 200 B
       

     The bending component K 200 B is primarily controlled by the diameter of rocker bearing portion  208 , and the distance between front and rear bores  418  and  420 . The rotating component K 200 R is primarily controlled by the length of the rocker shaft lever  300  and by the distance between the moveable axis  210  and fixed axis  422 . It is desirable to design the rotating component K 200 R such that it is greater than or equal to the bending component K 200 B. 
     The length of the rocker shaft lever  300  and the relative distances between the centerline  612 , moveable axis  210  and fixed axis  422  creates an advantageous lever ratio (i.e., greater than 1, preferably greater than 3 and more preferably greater than 5). Specifically, in this exemplary embodiment, this lever ratio (LR) is defined as the ratio of (1) the shortest distance between the line of action of the force (F 600 ) applied to the HLA  600  by rocker shaft lever  300  and the fixed axis  422  to (2) the shortest distance between the line of action of the force (F 200 ) applied to the rocker shaft assembly  200  by the rocker  162  and fixed axis  422 . 
     As the lever ratio increases above 1, it reduces the force from the rocker  162  onto the HLA assembly  600  (applied through rocker shaft lever  300 ), which increases the rotating component stiffness K 200 R relative to the HLA assembly stiffness K 600  by approximately the square of the lever ratio in accordance with the following equations:
 
 K 600 =F 600 /D 600  1)
 
 K 200 =F 200 /D 200  2)
 
 K 200 R=F 200 /D 200 R   3)
 
 K 200 B=F 200 /D 200 B   4)
 
1 /K 200=1 /K 200 R+ 1 /K 200 B   5)
 
 D 200 =D 200 R+D 200 B   6)
 
 D 600 =F 600 /K 600  7)
 
 F 600 =F 200 /LR   8)
 
 D 600 =F 200/( K 600 *LR )  9)
 
 D 200 R=D 600 /LR   10)
 
 D 200 R=F 200/( K 600 *LR*LR )  11)
 
 K 200 R=K 600 *LR*LR   12)
 
     If the preferable lever ratio (LR) of approximately 10 to 1 is used, the force (F 600 ) experienced by the plunger  630  of the HLA assembly  600  is only approximately one-tenth ( 1/10) of the force (F 200 ) experienced by the rocker shaft assembly  200  (as described in equation 8). At the same time, the deflection (D 600 ) in the general direction of axis  612  of the plunger  630  (due to the lever ratio of 10 to 1) is approximately 10 times the consequent deflection (D 200 R) in the general direction of axis  612  of the rocker shaft assembly  200  (as described in equation 10). 
     The overall result is that the lever ratio (LR) creates an effective increase in the rotating component (K 200 R) of the overall stiffness (K 200 ) of the rocker shaft assembly  200  compared to the stiffness (K 600 ) of the HLA assembly  600  that is approximately equal to the square of the lever ratio (as described in equation 12). One of the reasons that the relationship of stiffness k 200 R to stiffness K 600  is approximately, rather than exactly, that of equation 12 is friction. For purposes herein, the term “approximately”, as it applies to said square of said lever ratio, shall mean within 25 percent (or more preferably within 10 percent) of the value of said squared lever ratio. That is, if a lever ratio of approximately 10 to 1 is used (the preferred lever ratio), the rotating component stiffness K 200 R is approximately 100 times the HLA assembly stiffness K 600 . More specifically the stiffness of the rotating component K 200 R is preferably equal to or greater than 75 times the HLA assembly stiffness K 600 . More preferably, the stiffness of the rotating component K 200 R is equal to or greater than 90 times the HLA assembly stiffness K 600 . 
     As described above, the HLA assembly  600  is positioned remotely from the valve train  152 , which includes the cam  161 , rocker  162  and crossover valves  132 / 134  as actuating elements. Therefore, the primary motion of the rocker shaft lever  300  and the primary motion of the HLA assembly  600  will not be subject to the high frequency motion experienced by the actuating elements of the valve train  152  (about four to six times faster than the valves of a conventional engine). That is, the primary motion of the rocker shaft lever  300  and HLA assembly  600  (for example, the motion which compensates for variations in valve lash due to slower phenomenon, like thermal expansion, wear, HLA oil leakage and the like) will be at a much lower frequency than the primary motion of the actuating elements of the valve train  152 . Accordingly, the mass of the rocker shaft lever  300  will not be constrained by the high frequency motion requirements of valve train  152 . Therefore, the rocker shaft lever  300  can be made very stiff and bulky. Additionally, the lever ratio of rocker shaft lever  300  can be made very large, i.e., a lever ratio of 3 or greater, preferably a lever ratio of 5 or greater and most preferably a lever ratio of 7 or greater. 
     It should be noted that the rocker shaft lever  300  and HLA assembly  600  will be subject to some high frequency vibration caused by the high frequency movements of the valve train. However, the displacement induced by this vibration will have a magnitude that is substantially less than the magnitude of the displacement of the components in the valve train, typically by an order of magnitude less. The primary motion of the rocker shaft lever  300  and HLA assembly  600  in their lash adjustment function will have a frequency substantially less than that of the actuation motion of the actuating elements of the valve train  152 . 
     Although the valve lash adjustment system  160  described herein operates in conjunction with outwardly opening valves of a split-cycle engine, it can be applied to the operation of any valve. More preferably, it can be applied to fast acting valves having a duration of actuation of approximately 3 ms and 180 degrees of crank angle, or less.
         Although the invention has been described by reference to specific embodiments, it should be understood that numerous changes may be made within the spirit and scope of the inventive concepts described. For example, the valve lash adjustment system described herein is not limited to a cam-driven system. Accordingly, it is intended that the invention not be limited to the described embodiments, but that it have the full scope defined by the language of the following claims.