Patent Publication Number: US-7721684-B2

Title: Internal combustion engine

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This application is a continuation of application Ser. No. 10/969,362, filed Oct. 19, 2004, now U.S. Pat. No. 7,150,259, which is a continuation-in-part of application Ser. No. 10/627,288, filed Jul. 25, 2003, now U.S. Pat. No. 7,121,235, which is a continuation-in-part of application Ser. No. 10/147,372, filed May 15, 2002, now U.S. Pat. No. 6,598,567, which is a continuation-in-part of application Ser. No. 10/136,780, filed May 1, 2002, now abandoned, priority from the filing of which is hereby claimed under 35 U.S.C. § 120 and the disclosures of which are hereby expressly incorporated by reference. 

   FIELD OF THE INVENTION 
   The present invention is directed generally to combustion engines and, more particularly, to combustion engines having a power transfer assembly for transferring an energy of combustion generated within the combustion engine externally of the combustion engine for use by a device requiring power. 
   BACKGROUND 
   As is well known in the art, an internal combustion engine is a machine for converting heat energy into mechanical work. In an internal combustion engine, a fuel-air mixture that has been introduced into a combustion chamber is compressed as a piston slides within the chamber. A high voltage for ignition is applied to a spark plug installed in the combustion chamber to generate an electric spark to ignite the fuel-air mixture. The resulting combustion pushes the piston downwardly within the chamber, thereby producing a force that is convertible to a rotary output through the use of a crankshaft. 
   The crankshaft of the engine rotates at a high rate. Due to uneven forces acting upon the crankshaft, the crankshaft often vibrates when rotated. To counteract the uneven forces acting upon the crankshaft, counterweights are often rigidly coupled to, or as more often the case, formed integrally with the crankshaft, to balance the uneven forces acting upon the crankshaft, thereby eliminating/reducing vibrations in the crankshaft. 
   Although somewhat effective, previously developed crankshaft balancing techniques as described above are not effective in balancing crankshafts which both rotate and orbit so as to have two axes of rotation. Rigidly attaching counterweights directly to a crankshaft which both rotates and orbits would only effectively balance the crankshaft&#39;s rotation about its own axis, and would result in increased unbalancing forces during the crankshaft&#39;s orbital movement. Thus, there exists a need for a power transfer assembly having a crankshaft which is balanced while both rotating and orbiting. 
   Further, crankshafts which both rotate and orbit having additional problems. The movement of the crankshaft in both a rotational and orbital manner tends to magnify misalignment issues of the rotating components of the engine. Therefore, there exits a need for a power transfer assembly capable of mitigating or absorbing misalignment of the internal rotating components, such as a crankshaft, of a combustion engine. 
   SUMMARY 
   This summary is provided to introduce a selection of concepts in a simplified form that are further described below in the Detailed Description. This summary is not intended to identify key features of the claimed subject matter, nor is it intended to be used as an aid in determining the scope of the claimed subject matter. 
   One embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion and an outdrive. The outdrive is adapted to transfer the rotating motion of the crankshaft to an external device requiring power. The outdrive is non-rigidly interfaced with the crankshaft such that the crankshaft is permitted to freely rotate relative to the outdrive about at least one axis and freely move linearly in at least one direction relative to the outdrive during operation. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion and an outdrive. The outdrive is adapted to transfer the rotating motion of the crankshaft to an external device requiring power. The power transfer assembly further includes an interface assembly for non-rigidly interfacing the crankshaft with the outdrive to facilitate the transfer of power between the crankshaft and the outdrive. The interface assembly permits the crankshaft to move freely relative to the outdrive in at least one linear direction. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion and an outdrive. The outdrive is adapted to transfer the rotating motion of the crankshaft to an external device requiring power. The power transfer further includes an interface assembly for non-rigidly interfacing the crankshaft with the outdrive to facilitate the transfer of power between the crankshaft and the outdrive. The interface assembly directly transfers a torque from the crankshaft to the outdrive while simultaneously impeding transfer of centrifugal forces from the crankshaft to the outdrive. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion while rotating about a first axis and orbiting about a second axis and an outdrive. The outdrive is adapted to transfer the rotating motion of the crankshaft to an external device requiring power. The power transfer assembly further includes an interface assembly for facilitating the transfer of power between the crankshaft and the outdrive. The interface assembly non-rigidly interfaces the crankshaft with the outdrive such that a centrifugal force present in the crankshaft from the orbiting of the crankshaft about the second axis is not transferred to the outdrive during rotation of the crankshaft. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion and an outdrive. The outdrive is non-rigidly interfaced with the crankshaft and adapted to transfer the rotating motion of the crankshaft to an external device requiring power. The power transfer assembly further includes a crankshaft counterweight rotatably coupled to the crankshaft for reducing vibrations in the crankshaft during operation. The crankshaft counterweight is freely moveable along a path substantially radially oriented relative to an axis of rotation of the outdrive. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion. The crankshaft is adapted to rotate about a center axis which in turn orbits about an orbit axis. The power transfer assembly further includes a direct outdrive interfaced with the crankshaft. An orbital movement of the crankshaft causes rotary motion of the direct outdrive at a rate substantially equal to a rate at which the crankshaft orbits about the orbit axis. The power transfer assembly further includes a reduced outdrive interfaced with the crankshaft. A rotary movement and the orbital movement of the crankshaft combine to drive the reduced outdrive at a reduced rate relative to the direct outdrive. 
   Another embodiment of a power transfer assembly formed in accordance with the present invention for transferring energy within a combustion engine externally of the engine is disclosed. The power transfer assembly includes a crankshaft adapted to convert reciprocating motion to rotating motion. The crankshaft is adapted to rotate about a center axis which in turn orbits about an orbit axis. The power transfer assembly includes a first direct outdrive and a second direct outdrive, each interfaced with the crankshaft at opposite ends of the crankshaft. An orbital movement of the crankshaft causes rotary motion of the first and second direct outdrives at a rate substantially equal to a rate at which the crankshaft orbits about the orbit axis. The power transfer assembly further includes a reduced outdrive interfaced with the crankshaft. A rotary movement and the orbital movement of the crankshaft combine to drive the reduced outdrive at a reduced rate relative to the first and second direct outdrives. 

   
     DESCRIPTION OF THE DRAWINGS 
     The foregoing aspects and many of the attendant advantages of this invention will become more readily appreciated as the same become better understood by reference to the following detailed description, when taken in conjunction with the accompanying drawings, wherein: 
       FIG. 1  is a perspective view of one embodiment of a reciprocating internal combustion engine formed in accordance with the present invention, showing an engine block and related components, such as a control plate housing and an intake manifold, attached thereto; 
       FIG. 2  is a top planar view of the internal combustion engine depicted in  FIG. 1 ; 
       FIG. 3  is a side planar view of the internal combustion engine depicted in  FIG. 1 ; 
       FIG. 4  is a top planar view of the internal combustion engine depicted in  FIG. 1 , with a portion of the engine block cut-away, showing a cross-sectional view of a reciprocating cylinder liner receiving an opposing pair of substantially stationary pistons; 
       FIG. 5  is an elevation view of one embodiment of one of the substantially stationary pistons shown in  FIG. 4 ; 
       FIG. 6  is a cross-sectional view of one embodiment of the reciprocating cylinder liner shown in  FIG. 4 ; 
       FIG. 7  is a fragmentary cross-sectional view of a portion of the reciprocating cylinder liner and related components shown in  FIG. 4 , illustrating the reciprocating cylinder liner as a compression portion of a thermodynamic cycle is initiated; 
       FIG. 8  is a fragmentary cross-sectional view of the reciprocating cylinder liner and related components shown in  FIG. 4 , illustrating the reciprocating cylinder liner in a top-dead-center (TDC) position with respect to the shown substantially stationary piston as the reciprocating cylinder liner transitions into an expansion portion of the thermodynamic cycle; 
       FIG. 9  is a fragmentary cross-sectional view of the reciprocating cylinder liner and related components shown in  FIG. 4 , illustrating the reciprocating cylinder liner as the cylinder liner transitions into a scavenging portion of the thermodynamic cycle, marked by the opening of a plurality of intake ports near a crown of the substantially stationary piston and the opening of an exhaust valve; 
       FIG. 10  is a fragmentary cross-sectional view of the reciprocating cylinder liner and related components shown in  FIG. 4 , illustrating the reciprocating cylinder liner in a bottom-dead-center (BDC) position with respect to the shown substantially stationary piston as the reciprocating cylinder liner undergoes scavenging with the intake ports fully open and the exhaust valve fully open; 
       FIG. 11  is a fragmentary cross-sectional view of the reciprocating internal combustion engine of  FIG. 1 , the cross-sectional cut taken substantially along the centerline of a crank-cam so as to be coplanar with the centerline of a first cylinder liner and passing perpendicularly though the centerline of a second cylinder liner oriented normal to the first cylinder liner; 
       FIG. 12  is a perspective view of one embodiment of the crank-cam shown in  FIG. 11  formed in accordance with the present invention; 
       FIG. 13  is a bottom view of the crank-cam shown in  FIG. 12 ; 
       FIG. 14  is an elevation view of the crank-cam shown in  FIG. 12 ; 
       FIG. 15  is a side view of the crank-cam shown in  FIG. 14 ; 
       FIG. 16  is a diagrammatic elevation view showing the linear and rotary motion of the crank-cam with attached first and second cylinder liners; showing the first vertically oriented cylinder liner in an fully extended position and the second horizontally oriented cylinder liner in a mid-stroke position, wherein the distance between a pair of crank journals has been exaggerated to better show the movement of the cylinder liners; 
       FIG. 17  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 16 ; 
       FIG. 18  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 17 ; wherein the crank-cam has rotated 30° about a first axis of rotation from the position depicted in  FIG. 17 , showing the first vertically oriented cylinder liner as the liner moves linearly downward and the second horizontally oriented cylinder liner as it moves linearly to the left; 
       FIG. 19  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 18 ; 
       FIG. 20  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 16 ; wherein the crank-cam has rotated 90° about the first axis of rotation from the position depicted in  FIG. 16 , showing the first vertically oriented cylinder liner in a mid-stroke position and the second horizontally oriented cylinder liner in a fully extended position; 
       FIG. 21  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 20 ; 
       FIG. 22  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 16 ; wherein the crank-cam has rotated 150° about the first axis of rotation from the position depicted in  FIG. 16 , showing the first vertically oriented cylinder liner as the liner moves linearly downward and the second horizontally oriented cylinder liner as it moves linearly to the right; 
       FIG. 23  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 22 ; 
       FIG. 24  is a diagrammatic elevation view showing the linear and rotary motion of a crank-cam with attached first and second cylinder liners; wherein the crank-cam has rotated 180° about a first axis of rotation from the position depicted in  FIG. 16 ; showing the first vertically oriented cylinder in a fully extending position and the second horizontally oriented cylinder liner in a mid-stroke position; 
       FIG. 25  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 24 ; 
       FIG. 26  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 16 ; wherein the crank-cam has rotated 210° about a first axis of rotation from the position depicted in  FIG. 16 , showing the first vertically oriented cylinder liner as the liner moves linearly upward and the second horizontally oriented cylinder liner as it moves linearly to the right; 
       FIG. 27  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 26 ; 
       FIG. 28  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 16 ; wherein the crank-cam has rotated 270° about the first axis of rotation from the position depicted in  FIG. 16 ; showing the first vertically oriented cylinder liner in a mid-stroke position and the second horizontally oriented cylinder liner in a fully extended position; 
       FIG. 29  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 28 ; 
       FIG. 30  is a diagrammatic elevation view of the crank-cam with attached first and second cylinder liners of  FIG. 16 ; wherein the crank-cam has rotated 360° about the first axis of rotation from the position depicted in  FIG. 16 , showing the first vertically oriented cylinder liner in a fully extend position and the second horizontally oriented cylinder liner in a mid-stroke position; 
       FIG. 31  is a diagrammatic side view of the crank-cam with attached first and second cylinder liners depicted in  FIG. 30 ; 
       FIG. 32  is an exploded view of a crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange, suitable for use with the illustrated embodiment of the present invention, wherein the outdrive gear is shown in cross-section and the outdrive reduction gear is shown with a partial cut-away; 
       FIG. 33  is a planar cross-sectional end view of the outdrive gear, outdrive reduction gear, power take-off flange, and crank-cam shown in  FIG. 32 , taken substantially through section  33 - 33  of  FIG. 32 ; 
       FIG. 34  is a planar end view of the crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange shown in  FIG. 32 , wherein the outdrive reduction gear has rotated 1/16 of a turn from its position depicted in  FIG. 32 ; 
       FIG. 35  is a planar end view of the crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange shown in  FIG. 32 , wherein the outdrive reduction gear has rotated ⅛ of a turn from its position depicted in  FIG. 32 ; 
       FIG. 36  is a planar end view of the crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange shown in  FIG. 32 , wherein the outdrive reduction gear has rotated ¼ of a turn from its position depicted in  FIG. 32 ; 
       FIG. 37  is a planar end view of the crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange shown in  FIG. 32 , wherein the outdrive reduction gear has rotated ⅜ of a turn from its position depicted in  FIG. 32 ; 
       FIG. 38  is a planar end view of the crank-cam, outdrive gear, outdrive reduction gear, and power take-off flange shown in  FIG. 32 , wherein the outdrive reduction gear has rotated ½ of a turn from its position depicted in  FIG. 32 ; 
       FIG. 39  is a planar end view of a direct outdrive and a gliding block formed in accordance with the present invention; 
       FIG. 40  is an exploded top view of the direct outdrive and the gliding block shown in  FIG. 39 ; 
       FIG. 41  is an exploded side view of the direct outdrive and the gliding block shown in  FIG. 39 , and in addition showing a direct outdrive adapter; 
       FIG. 42  is a planar end view of the direct outdrive, gliding block, and direct outdrive adapter shown in  FIG. 41 ; 
       FIG. 43  is a planar end view of the direct outdrive, gliding block, and outdrive adapter shown in  FIG. 42 , where the direct outdrive has rotated 90° from its position depicted in  FIG. 42 ; 
       FIG. 44  is a planar end view of the direct outdrive, gliding block, and outdrive adapter shown in  FIG. 42 , where the direct outdrive has rotated 180° from its position depicted in  FIG. 42 ; 
       FIG. 45  is a planar end view of the direct outdrive, gliding block, and outdrive adapter shown in  FIG. 42 , where the direct outdrive has rotated 270° from its position depicted in  FIG. 42 ; 
       FIG. 46  is a partial cross-sectional view of an alternate embodiment of a power transfer assembly formed in accordance with the present invention and suitable for use with the reciprocating internal combustion engine of  FIGS. 1-45 ; 
       FIG. 47  is an end view of a direct outdrive assembly shown in  FIG. 46 ; 
       FIG. 48  is an exploded elevation view of the direct outdrive assembly shown in  FIG. 47 ; 
       FIG. 49  is an exploded view of a portion of the power transfer assembly depicted in  FIG. 46 ; 
       FIG. 50  is an end view of a direct outdrive assembly portion of a direct and reduction outdrive assembly shown in  FIG. 46 ; 
       FIG. 51  is an exploded elevation view of the direct outdrive assembly portion shown in  FIG. 50 ; 
       FIGS. 52A-52H  are cross-sectional views taken substantially through Section  52 A- 52 A of  FIG. 46 , depicting the crankshaft assembly and direct outdrive assembly sequentially as the crankshaft assembly and direct outdrive assembly rotate through one cycle; 
       FIGS. 53A-53H  are cross-sectional views taken substantially through Section  53 A- 53 A of  FIG. 46 , depicting the crankshaft assembly and a direct and reduction outdrive assembly sequentially as the crankshaft assembly and the direct and reduction outdrive assembly rotate through one cycle; 
       FIG. 54  is an alternate embodiment of the power transfer assembly depicted in  FIG. 46 , wherein a pair of direct outdrives are reduced in size to permit enlargement of a pair of counterbalance weights; 
       FIG. 55  is an alternate embodiment of the direct outdrive assembly shown in  FIGS. 50 and 51 ; and 
       FIG. 56  is a cross-section view of the direct outdrive assembly of  FIG. 55  taken substantially through Section  56 - 56  of  FIG. 55 . 
   

   DETAILED DESCRIPTION 
     FIGS. 1-45  illustrate one embodiment of a reciprocating internal combustion engine  1010  formed in accordance with the present invention. The engine  1010  is unlike conventional reciprocating internal combustion engines, in that the engine  1010  reciprocates two cylinder liners  1014   a  and  1014   b , orthogonally oriented relative to one another, between opposing pairs of “substantially stationary” pistons  1012   a  and  1012   b , and  1012   c  and  1012   d  respectively. As used within this detailed description, the phrase “substantially stationary” is intended to mean a part, that although may be capable of some movement, does not move in accordance with a crankshaft or analogous component of an engine, as does a piston, camshaft, connecting rod, or valve of a conventional engine. In other words, a substantially stationary part&#39;s movement is separate and independently actuatable relative to the crankshaft or analogous component of an engine. In conventional reciprocating internal combustion engines, the pistons are reciprocated within stationary cylinders. 
   In the embodiment illustrated in  FIGS. 1-45 , many of the components are identical to one another, such as the pistons  1012   a ,  1012   b ,  1012   c , and  1012   d  and each of the two cylinder liners  1014   a  and  1014   b . Therefore, a numbering scheme has been adopted in which components of identical structure are assigned a common reference numeral followed by a selected letter to distinguish them from their identical counterpart. Where the context permits, reference in the following description to an element of one component having an identical counterpart shall be understood as also referring to the corresponding element of the identical counterpart. 
   Referring now to  FIGS. 1-3 , an engine block  1013  and other related external components of one illustrated embodiment formed in accordance with the present invention will be discussed. The engine block  1013  is suitably an octagonal block structure having an upper planar end surface  1146  opposite a lower planar end surface  1148  with internal cavities for housing the pistons, cylinders, and other related components there between. The engine block  1013  is formed from a rigid material, such as steel, cast iron, or aluminum, by techniques well known in the art, such as machining and/or casting. Fastened to the sidewalls of the engine block  1013  are two intake manifolds  1138  and four square mounting plates  1136 . Coupled to each of the mounting plates  1136  is a housing mounting plate  1144 , upon each of which is coupled a control plate housing  1320 . 
   Referring now to  FIGS. 1 and 4 , the housing mounting plate  1144  will be described. The housing mounting plate  1144  serves as an insulator, impeding the transfer of heat generated in the engine block  1013  to the various components of a compression ratio and power setting control system  1300 , which will be described in further detail below. To impede heat transfer, the housing mounting plate  1144  contains an inner cavity  1324 . The inner cavity  1324  impedes heat transfer by limiting the contact between components of the compression ratio and power setting control system  1300  and the mounting plate  1136 . Further, the housing mounting plate  1144  includes four cooling ports  1326  in fluid communication with the inner cavity  1324  and the outer environment, to allow heated air to exchange with exterior cool air. 
   Referring again to  FIGS. 1-3 , protruding from the control plate housings  1320  are the distal ends of each of the pistons  1012  and upper chamber piping  1312  associated with the compression ratio and power setting control system  1300 . Protruding from the housing mating plate  1144  is lower chamber piping  1314  also associated with the compression ratio and power setting control system  1300 . Located above or below the control plate housing  1320 , as the case may be, is an exhaust port  1142 . The exhaust ports  1142  are in fluid communication with the exhaust gas passages  1037  (see  FIG. 10 ) located internally in the engine block  1013 , and allow the discharge of products of combustion generated in the combustion chambers of the engine  1010  to the atmosphere. Preferably, well known exhaust gas collection, treatment, and/or muffler systems (not shown) are coupled in fluid communication with the exhaust ports  1142 . Each intake manifold  1138  includes two intake ports  1140 . Preferably coupled to each intake port  1140  are well-known intake systems that may include such components as a carburetor and/or a filter. 
   Referring to  FIG. 4  and focusing mainly now on the internal components of the internal combustion engine  1010 , the engine  1010  includes two double cylinder liners  1014   a  and  1014   b , each of which houses two substantially stationary opposing pistons  1012   a  and  1012   b  and  1012   c  and  1012   d , respectively, in opposite ends of the cylinder liners  1014   a  and  1014   b . The cylinder liners  1014   a  and  1014   b  are perpendicularly and offset mounted relative to one another within the engine block  1013 . The cylinder liners  1014   a  and  1014   b  alternately reciprocate between a first extended position and a second extended position. More specifically, with reference to cylinder liner  1014   a , the cylinder liner  1014   a  reciprocates between a first extended position wherein the cylinder liner  1014   a  is at a top-dead-center (TDC) position relative to a first piston  1012   b  and a bottom-dead-center (BDC) position relative to a second piston  1012   a , as shown in  FIG. 4 , and a second extended position, where the cylinder liner  1014   a  is at a BDC position relative to the first piston  1012   b  and a TDC position relative to the second opposing piston  1012   a . The second cylinder liner  1014   b  similarly reciprocates between a first extended position and a second extended position. However, the second cylinder liner  1014   b  reciprocates 180° out of phase of the first cylinder liner  1014   a  so that when the first cylinder liner  1014   a  is in extended position, the second cylinder liner  1014   b  is in a mid-stroke position. The cylinder liners  1014  are coupled to one another by a crankshaft, which will be referred to as a crank-cam  1016  for the purposes of this detail description. The crank-cam  1016  converts the linear motion of the cylinder liners  1014  to rotary motion, as will be discussed in further detail below. 
   Referring to  FIG. 5 , the physical structure of one of the four substantially stationary pistons  1012  formed in accordance with the present invention will now be described. Inasmuch as the pistons  1012  are substantially identical to one another, reference to the piston  1012   a , illustrated in  FIG. 5 , shall be understood as also referring to the corresponding other three pistons  1012   b ,  1012   c , and  1012   d  (see  FIG. 4 ) where context permits. The piston  1012   a  is a hollowed, cylindrical plunger having a piston head  1018  concentrically and perpendicularly mounted to a shaft  1020 . Both the piston head  1018  and shaft  1020  have aligned internal bores, forming a channel  1022  running axially through the center of the piston  1012 . The channel  1022  allows a substantial reduction in the weight of the piston  1012 , while also permitting access to the spark plug  1024  and/or a fuel injector (not shown) disposed within the piston head  1018 . The pistons  1012  contain a spark plug or injector hole  1023  for the mounting of a spark plug  1024  and/or fuel injector therein. 
   Circumferentially mounted on the piston head  1018  are two compression rings  1030 . As is well known in the art, the compression rings  1030  prevent the blow-by of combustion gases and products past the piston head  1018 , mainly during the compression and expansion portions of the thermodynamic cycle. Although not shown, the piston head  1018  may also include an oil control ring, as is well known in the art. In proximity to the compression rings  1030 , the diameter of the piston head  1018  is substantially equal to the diameter of the cylinder liner  1014 . The diameter of the piston head  1018  may be tapered thereafter along the length of the piston head  1018 , resulting in a portion of the piston head  1018  spaced from the compression rings having a relatively smaller diameter. 
   Circumferentially mounted on the shaft  1020  is a compression ratio control plate  1026 . The compression ratio control plate  1026  is adaptable to receive pressurized control fluid on the upper and lower annular surfaces  1025  and  1027  of the plate  1026 . By selectively providing a pressure differential across the annular surfaces  1025  and  1027 , the axial position of the piston  1012   a  may be adjusted relative to the engine block to allow the power setting and compression ratio of the engine to be adjusted, as will be described in greater detail below. Two oil control rings  1028  are circumferentially mounted on the compression ratio control plate  1026  to prevent the leakage of any control fluid thereby. 
   Referring to  FIG. 6 , reciprocating double cylinder liner  1014   a , which operates in conjunction with two of the above-described substantially stationary pistons  1012 , will now be described. Inasmuch as the double cylinder liners  1014  are substantially identical to one another, reference to the cylinder liner  1014   a  illustrated in  FIG. 6  shall be understood as also referring to the other cylinder liner  1014   b  (see  FIG. 4 ), where context permits. The double cylinder liner  1014   a  is a generally elongate cylindrical structure having a first axially aligned bore concentrically formed in an upper distal end of the cylinder liner  1014   a , thereby forming a first cylinder  1032   a  for reciprocatingly receiving a piston  1012   a  (see  FIG. 4 ). Located on an opposite lower distal end of the cylinder liner  1014   a  is a second concentrically formed, axially aligned bore in the cylinder liner  1014   a , thereby forming a second cylinder  1032   b  for reciprocatingly receiving a second piston  1012   b  (see  FIG. 4 ). The cylinders  1032   a  and  1032   b  are shaped and sized to receive the pistons  1012   a  and  1012   b  in a clearance fit relationship, as is well known in the art. 
   Referring now to  FIGS. 4 ,  6 , and  7 , at the inner or bottom ends of the cylinders  1032  are exhaust valve seats  1034 . The exhaust valve seats  1034  are formed by well-known techniques in the art to receive an exhaust valve there within. In fluid communication with the exhaust valve seats  1034  are four exhaust gas passages  1036  for discharging exhaust gases from the cylinders  1032 . Centrally bored through the cylinder liner  1014   a  is a valve stem bore  1038 . The valve stem bore  1038  is sized to receive a stem of the exhaust valve  1052 . In communication with the valve stem bore  1038  is a valve spring housing  1040 . The valve stem housing  1040  is sized and configured to house a spring for biasing the exhaust valve in the closed position. In communication with the valve spring housing  1040  is a crank-cam housing  1042 . The crank-cam housing  1042  is sized and configured to house the crank-cam  1016  and allow its rotation therewithin. 
   Referring now to  FIGS. 6 and 11 , the crank-cam housing  1042  is formed by a cylindrically shaped bore  1150  perpendicularly passing through the cylinder liner  1014   a  at a location equidistant from the ends of the cylinder liner. The radius of the bore  1150  is substantially equal to the distance measured from the centerline of the crank-cam  1016  to an outer surface of a crank-cam  1016  crank journal  1072 . A radius of this dimension allows the crank journal to rotate freely within the bore  1150  of the crank-cam housing  1042  during operation. The diameter of the bore  1150  is stepped suddenly outward in the center of the bore  1150  to form a lobe clearance bore  1152 . The radius of the lobe clearance bore  1152  is equal to or greater than a distance measured from a centerline of the crank-cam to the distal end or peak of the lobe  1054  of the crank-cam  1016 . A radius of this dimension provides sufficient clearance for the lobe  1054  to rotate freely within the crank-cam housing  1042 . 
   Located on opposite distal ends of the cylinder liner  1014   a  are annular precompression plates  1044 . The annular precompression plates  1044  are utilized to compress and deliver pressurized combustion gases to the cylinders  1032 , as will be discussed in more detail below. In proximity to the annular precompression plates  1044  are intake ports  1046 . In the illustrated embodiment, the intake ports  1046  are spaced circumferentially about the cylinders  1032  at 60° intervals; however, it should be apparent to one skilled in the art that other configurations are suitable. The intake ports  1046  allow the entry of combustion gases into the cylinders  1032  during operation for scavenging and charging of the cylinders  1032 . Located on the inner and outer surfaces of the annular precompression plates are inner and outer combustion gas/oil seals  1048 . The seals  1048  prevent the passage of fluids thereby as will be described in more detail below. 
   Referring now to  FIG. 7 , in light of the above description of the reciprocating double cylinder liners  1014  and the substantially stationary pistons  1012 , the relationship of these and related components to one another during significant events in a thermodynamic cycle will now be discussed. The illustrated embodiment of the reciprocating internal combustion engine  1010  of the present invention operates on a two-stroke cycle. Therefore, for every revolution of the crank-cam  1016 , each piston  1012  completes the thermodynamic cycle in two strokes, a single stroke defined by movement of the cylinder liner  1014  from a TDC position to a BDC position (or vice versa) relative to the substantially stationary pistons  1012  contained within the cylinder liners  1014 . Therefore, every stroke of the cylinder liner  1014  is either a power stroke, also known as an expansion stroke, or a compression stroke relative to each piston  1012 . This requires the intake and exhaust functions, i.e., scavenging, to occur rapidly at the end of each power stroke and before the succeeding compression stroke. In the illustrated embodiment, each piston  1012  undergoes one power stroke for each revolution of the crank-cam  1016 , resulting in twice as many power strokes as in a similarly designed four-stroke cycle engine for a given RPM. 
   Still referring to  FIG. 7 , the cylinder liner  1014  is depicted at the commencement of the compression portion of the thermodynamic cycle. More specifically, the cylinder liner  1014  is depicted as it moves upward from the cylinder liner&#39;s BDC position toward the piston  1012 . As cylinder liner  1014  moves upward, the piston  1012  completely covers the intake ports  1046 , thereby sealing off the cylinder  1032 . In the depicted position, an exhaust lobe  1054  on the crank-cam  1016  is oriented just as the valve stem  1066  comes off of the exhaust lobe  1054 , thereby allowing a valve spring  1056  to bias an exhaust valve  1052  into a closed position. In the closed position, the exhaust valve  1052  sealingly engages an exhaust valve seat  1034  in the cylinder liner  1014 , thereby preventing the discharge of any combustion gases from the cylinder  1032 . Configured as described, the combustion gases are sealingly contained within a combustion chamber  1033 , defined by the side and bottom peripheral walls of the cylinder  1032  and the end surface, or crown  1019  of the piston head  1018 . 
   As the cylinder liner continues to approach the piston, departing from its BDC position and approaching its TDC position relative to the piston  1012 , the volume of the combustion chamber  1033  is accordingly decreased, thereby compressing the combustion gases contained therewithin. Referring now to  FIG. 8 , when, or just prior to arrival of the cylinder liner  1014  at its TDC position respective to the piston  1012 , a high voltage spark  1058  is discharged from the spark plug  1024  (see  FIG. 5 ) by well-known means, thereby igniting the combustion gases. As the combustion gases burn, the resulting products of combustion expand, driving the cylinder liner  1014  away from the piston  1012 . Referring now to  FIG. 9 , the expansion of the products of combustion continues to drive the cylinder liner  1014  down and away from the piston  1012 , until the point in the cycle wherein the exhaust valve  1052  is displaced from its seat  1034  and the intake ports  1046  are uncovered, thus initiating the scavenging of the products of combustion from the combustion chamber  1033 . 
   However, prior to scavenging the products of combustion from the combustion chamber  1033 , a new volume of combustion gases is pressurized to aid in scavenging of the combustion chamber  1033 . In the illustrated embodiment of the present invention, this is accomplished by the sweeping of the annular precompression plates  1044  through an intake chamber  1064 . More specifically, as the cylinder liner  1014  travels upward from the position shown in  FIG. 7  to the position shown in  FIG. 8 , the annular precompression plate  1044  is forced to sweep through the cylindrically-shaped intake chamber  1064 . As the precompression plate  1044  sweeps upward through the intake chamber  1064 , a vacuum is created within the intake chamber  1064 , which draws new combustion gases into the intake chamber  1064 . A well-known one-way reed check valve (not shown) allows the flow of the combustion gases into the intake chamber  1064 , while preventing the passage of any combustion gases or products of combustion out of the intake chamber  1064 . 
   As the cylinder liner  1014  travels downward from the position shown in  FIG. 8  to the position shown in  FIG. 9 , i.e., from a TDC position to a BDC position, the intake chamber  1064  is a sealed pressure vessel as the intake ports  1046  are sealed off by the piston  1012  and the one-way reed check valves prevent the discharge of combustion gases out the intake chamber  1064 . As the precompression plate  1044  sweeps downward through the intake chamber  1064 , the combustion gases contained in the intake chamber  1064  are compressed until released into the combustion chamber  1033  by the uncovering of the intake ports  1046 . 
   The intake chamber  1064  preferably contains a volume greater than the maximum displacement of the combustion chamber  1033 . In the illustrated embodiment, the intake chamber  1064  is three times larger than the maximum displacement of the combustion chamber, although it should be apparent to one skilled in the art that other ratios of intake chamber volume to maximum combustion chamber volume are suitable for use with the present invention, such as low as 1:1 and up to 3:1 or higher. As a result of the relatively greater volume of the intake chamber  1064  relative to the combustion chamber  1033 , combustion gases may be provided at an elevated pressure. Thus, by selecting the relative size of the intake chamber  1064 , combustion gases at elevated pressures similar to those reached in a super-charged or turbo-charged conventional engine may be achieved. The pressurization of the combustion gases occurs even at low RPMs, unlike conventional super-charged or turbo-charged engines, which typically are unable to provide sufficient pressurization of the combustion gases at low RPM, resulting in a lag in engine performance as the engine reaches an elevated RPM able to provide sufficiently pressurized combustion gases. 
   Scavenging of the combustion chamber  1033  commences at the end of the power stroke. The end of the power stroke is marked by the opening of the intake ports  1046  and the exhaust valve  1052 . This occurs, as depicted in  FIG. 9 , as the cylinder liner  1014  moves down and away from the substantially stationary piston  1012  to the point that the intake ports  1046  are initially uncovered and the exhaust valve  1052  is initially lifted from its seat  1034 . As the intake ports  1046  are initially uncovered, the pressurized combustion gases contained within the intake chamber  1064  below the precompression plate  1044  are released into the combustion chamber  1033 . At approximately the same time, the exhaust valve  1052  is initially lifted off the valve seat  1034  as the lobe  1054  of the crank-cam  1016  engages the valve stem  1066 , thereby disposing the exhaust valve  1052  toward the substantially stationary piston  1012 . Thus, the products of combustion contained in the combustion chamber  1033  begin to be swept from the combustion chamber  1033  as the pressurized combustion gases contained in the intake chamber  1064  are released from the intake chamber  1064  through the intake ports  1046  and through the combustion chamber  1033 . The entrance of the pressurized combustion gases into the combustion chamber  1033  forces the products of combustion out the exhaust gas passageways  1036  in the cylinder liner  1014  as they align with the exhaust gas passageways  1037  located in the engine block  1013 . 
   The exhaust gas passageways  1037  are centrally located in the engine block  1013  and are alternately aligned depending upon the position of the cylinder liner  1014 , in fluid communication with a first pair of exhaust gas passageways  1036   a  and a second pair of exhaust gas passageways  1036   b  in the cylinder liners  1014 . More specifically, when the cylinder liner  1014  is at a BDC position with respect to a first piston  1012   a , the first pair of exhaust gas passageways  1036   a  associated with the first piston  1012   a  are in fluid communication with the exhaust gas passageways  1037  in the engine block  1013 . When the cylinder liner moves to a BDC position with respect to a second piston opposing the first piston, the second pair of exhaust gas passageways  1036   b  associated with the second piston will be in fluid communication with the exhaust gas passageways  1037  in the engine block  1013 . 
   Returning now to the operation of the engine, the cylinder liner  1014  continues to move away from the substantially stationary piston  1012   a  until the cylinder liner  1014  reaches BDC. At BDC, as depicted in  FIG. 10 , the intake ports  1046  and exhaust valve  1052  are fully open. At this point, the pressurized combustion gases are flowing into the combustion chamber  1033  at a high rate, thus purging the combustion chamber  1033  of the products of combustion and recharging the combustion chamber  1033  with fresh combustion gases. As the crank-cam  1016  continues to rotate clockwise past the BDC position, the exhaust valve  1052  retracts into a closed position as the lobe  1054  disengages from the valve stem  1066  and the cylinder liner  1014  moves toward the substantially stationary piston  1012 , thereby closing off the intake ports  1046 . Thus, the combustion chamber  1033  is completely sealed and the combustion gases contained therewithin begin to be compressed, thus returning the cycle to the position depicted in  FIG. 7 . 
   Referring to  FIGS. 12-15 , the crank-cam  1016  of the illustrated embodiment will now be described in further detail. The crank-cam  1016  serves both the functions of a crankshaft and a camshaft in a conventional reciprocating internal combustion engine. The crank-cam  16  includes three circular crank webs  1070 , two crank journals  1072   a  and  1072   b , and two crank-cam lobes  1054 . The crank-cam  1016  may be of steel or other suitably rigid material, forged in one piece, or may be built up, such as by shrink-fitting separately forged crank journals  1072  to cast crank webs  1070 . Although the crank webs  1070  are concentrically aligned relative to one another, the crank journals  1072  are offset relative to one another by a distance equal to one half of the stroke length and are also offset relative to the centerline  1074  of the crank webs  1070 . 
   Referring now to FIGS.  4  and  12 - 15 , the crank journals  1072   a  and  1072   b  are disposed relative to one another so that when a first cylinder liner  1014   a  is in a TDC relationship relative to one piston  1012   b  and at a BDC relationship to a second opposing piston  1012   a , the second cylinder liner  1014   b  is equidistant from its opposing pistons  1012   c  and  1012   d . Likewise, the crank-cam lobes  1054  of each respective crank journal  1072  face in opposite directions, so that when the first crank-cam lobe  1054   a  has positioned an exhaust valve  1052  in its fully open position relative to a piston  1012   a , the other crank-cam lobe  1054   b  is equidistant from the opposing substantially stationary pistons  1012   c  and  1012   d , and therefore does not engage the valve stems of either exhaust valve, thus placing the respective exhaust valves in a closed position. 
   As should be apparent to one skilled in the art, the force to compress the combustion gases associated with a first piston  1012   a  is provided by the expansion of the gases related to the opposing piston  1012   b . Therefore, as should be apparent to one skilled in the art, the force exerted upon the crank journal  1072   a  is a resultant force of an expansion force generated by the expansion of the combustion gases minus a compression force required to compress the combustion gases related to the opposing piston. Further, inasmuch as the compression force and the expansion force are collinear, a moment is not created upon the crank-cam  1016  by the simultaneous application of the expansion and compression forces. Thus, the crank-cam  1016  of the present invention may be reduced in size relative to a crankshaft of a conventional engine that does not counter the expansion force with a collinear compression force. 
   Referring now to  FIGS. 12-15  and  16 - 31 , the relationship between the cylinder liners  1014   a  and  1014   b  relative to the crank-cam  1016  during operation will now be described. Referring to  FIGS. 16 and 17 , wherein  FIG. 17  is a side view of the components depicted in  FIG. 16 , a first cylinder liner  1014   a  is mounted vertically on a first crank journal  1072   a . A second cylinder liner  1014   b  is perpendicularly, and thus horizontally, mounted relative to the first cylinder liner  1014   a  on a second crank journal  1072   b . The first cylinder liner  1014   a  is restricted to a vertical reciprocating path of travel by the engine block represented by the line identified by the reference numeral  1100 . Likewise, the second cylinder liner  1014   b  is restricted by the engine block to a horizontal-reciprocating path of travel represented by the line identified by the reference numeral  1098 . 
   The reciprocating linear motion of the cylinder liners  1014   a  and  1014   b  is translated into rotary motion via the crank-cam  1016 . More specifically, the crank-cam  1016  rotates on two axes of rotation. The first axis of rotation  1074  is about the centerline of the crank-cam  1016 . More specifically, the first axis of rotation  1074  is defined by a line coplanar, parallel, and equidistant from the centerline  1076   a  and  1076   b  of each crank journal  1072   a  and  1072   b . During operation, the crank-cam  1016  rotates about the first axis of rotation  1074 , while the first axis of rotation  1074  is further rotated in a circular orbit  1080  around a second axis of rotation  1078 . The second axis of rotation  1078  is defined as a line normal to both the centerline of the first cylinder liner  1014   a  and the second cylinder liner  1014   b  that bisects the midpoint of the strokes of each cylinder liner  1014   a  and  1014   b . The radius of the circular orbit  1080  from the second axis of rotation  1078  is equal to one-quarter of the stroke length. 
   Still referring to  FIGS. 16 and 17 , cylinder liner  1014   a  is depicted in an extended position, where the cylinder liner  1014   a  is in a TDC and a BDC position relative to its two opposing pistons, while cylinder liner  1014   b  is depicted in a midpoint position, where the cylinder liner  1014   b  is equidistant from its respective opposing pistons. In this configuration, the second axis of rotation  1078  is collinear with the centerline of the crank journal  1072   b  and bisects the midpoint of the stroke length of cylinder liner  1014   b . As the crank-cam rotates clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  move linearly to the left along the horizontal path of travel  1098  of the cylinder liner  1014   b . Likewise, crank-journal  1072   a  and its related cylinder liner  1014   a  move linearly downward along the vertical path of travel  100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 18 and 19 . 
   Referring to  FIGS. 18 and 19 , the crank-cam with attached cylinder liners  1014   a  and  1014   b  are shown after the crank-cam has rotated 30° about the first axis of rotation  1074 . Thus, cylinder liner  1014   a  is depicted as it moves linearly downward and away from its extended position depicted in  FIGS. 16 and 17  and cylinder liner  1014   b  is depicted as it travels left from the midpoint position depicted in  FIGS. 18 and 19 . As the crank-cam rotates clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  move linearly to the left along the horizontal path of travel  1098  of the cylinder liner  1014   b . Likewise, crank-journal  1072   a  and its related cylinder liner  1014   a  move linearly downward along the vertical path of travel  1100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 20 and 21 . 
   Referring now to  FIGS. 20 and 21 , the crank-cam with attached cylinder liners  1014   a  and  1014   b  are shown after the crank-cam has rotated 90° about the first axis of rotation  1074 . Thus, cylinder liner  1014   b  is depicted in an extended position relative to its two opposing pistons, while cylinder liner  1014   a  is depicted in a midpoint position, where the cylinder liner  1014   a  is equidistant from its respective opposing pistons. In this configuration, the second axis of rotation  1078  is collinear with the centerline  1076   a  of the crank journal  1072   a  and bisects the midpoint of the stroke length of cylinder liner  1014   a . As the crank-cam continues to rotate clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  change direction and now move linearly to the right along the horizontal path of travel  1098  of the cylinder liner  1014   b . Crank-journal  1072   a  and its related cylinder liner  1014   a  continue to move linearly downward along the vertical path of travel  1100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 22 and 23 . 
   Referring now to  FIGS. 22 and 23 , the crank-cam with attached cylinder liners  1014   a  and  1014   b  are shown after the crank-cam has rotated 150° about the first axis of rotation  1074 . Thus, cylinder liner  1014   a  is depicted as it moves linearly downward from its midway position depicted in  FIGS. 20 and 21  and cylinder liner  1014   b  is shown as the cylinder liner  1014   b  travels right from its extended position depicted in  FIGS. 20 and 21 . As the crank-cam rotates clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counterclockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  moves linearly to the right along the horizontal path of travel  1098  of the cylinder liner  1014   b  to its midpoint position. Likewise, crank-journal  1072   a  and its related cylinder liner  1014   a  move linearly downward along the vertical path of travel  1100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 24 and 25 . 
   Referring to  FIGS. 24 and 25 , cylinder liner  1014   a  is depicted in a extended position, where the cylinder liner  1014   a  is in a TDC and BDC position relative to its two opposing pistons, while cylinder liner  1014   b  is depicted in a midpoint position, where the cylinder liner  1014   b  is equidistant from its respective opposing pistons. In this configuration, the second axis of rotation  1078  is collinear with the centerline of the crank journal of the cylinder liner  1014   b  and bisects the midpoint of the stroke length of cylinder liner  1014   b . As the crank-cam rotates clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  move linearly to the right along the horizontal path of travel  1098  of the cylinder liner  1014   b . Likewise, crank-journal  1072   a  and its related cylinder liner  1014   a  move linearly upward along the vertical path of travel  100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 26 and 27 . 
   Referring to  FIGS. 26 and 27 , the crank-cam with attached cylinder liners  1014   a  and  1014   b  are shown after the crank-cam has rotated 210° about the first axis of rotation  1074 . Thus, cylinder liner  1014   a  is depicted as it moves linearly upward and away from its extended position depicted in  FIGS. 24 and 25  and cylinder liner  1014   b  is depicted as it travels right from the equidistant position depicted in  FIGS. 24 and 25 . As the crank-cam rotates clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  move linearly to the right along the horizontal path of travel  1098  of the cylinder liner  1014   b . Likewise, crank-journal  1072   a  and its related cylinder liner  1014   a  move linearly upward along the vertical path of travel  1100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 28 and 29 . 
   Referring now to  FIGS. 28 and 29 , the crank-cam with attached cylinder liners  1014   a  and  1014   b  are shown after the crank-cam has rotated 270° about the first axis of rotation  1074 . Thus, cylinder liner  1014   b  is depicted in an extended position relative to its two opposing pistons, while cylinder liner  1014   a  is depicted in a midpoint position, where the cylinder liner  1014   b  is equidistant from its respective opposing pistons. In this configuration, the second axis of rotation  1078  is collinear with the centerline of the crank journal  1072   b  and bisects the midpoint of the stroke length of cylinder liner  1014   b . As the crank-cam continues to rotate clockwise about the first axis of rotation  1074  while the first axis of rotation  1074  simultaneously rotates counter clockwise along the circular orbit  1080  centered around the second axis of rotation  1078 , crank-journal  1072   b  and its related cylinder liner  1014   b  change direction and now move linearly to the left along the horizontal path of travel  1098  of the cylinder liner  1014   b . Crank-journal  1072   a  and its related cylinder liner  1014   a  continue to move linearly upward along the vertical path of travel  1100  of its related cylinder liner  1014   a  to the configuration shown in  FIGS. 30 and 31 , thus returning the engine to the configuration depicted in  FIGS. 16 and 17 , marking the completion of a single thermodynamic cycle relative to each piston. 
   Referring now to  FIG. 11 , the interrelationship between the crank-cam  1016  and the cylinder liners  1014   a  and  1014   b  will now be described in further detail.  FIG. 11  depicts a fragmentary cross-section of a reciprocating internal combustion engine  1010  formed in accordance with the present invention. The cross-section is taken substantially along the longitudinal length of the crank-cam  1016 . With the cross-section taken as such, the vertically oriented-cylinder liner  1014   a  is sectioned along the centerline of the cylinder liner  1014   a . Inasmuch as cylinder liner  1014   b  is orientated normal to cylinder liner  1014   a , and thus in a horizontal orientation, the cross-section passes laterally through cylinder liner  1014   b  midway between the ends of the cylinder liner  1014   b . Cylinder liner  1014   a  is shown in a BDC configuration relative to piston  1012   a  (not shown) and in a TDC relationship relative to piston  1012   b.    
   Cylinder liner  1014   b  is shown equidistant from its opposing pistons. With the crank-cam  1016  configured as such, the lobe  1054   a  associated with the crank journal  1072   a  has engaged the valve stem  1066   a  of the exhaust valve  1052  associated with piston  1012   a , lifting the valve  1052  off of its seat  1034 . The lobe  1054   b  associated with the crank journal  1072   b  of cylinder liner  1014   b  is shown equidistant between the valve stems of the opposing substantially stationary pistons. Inasmuch as cylinder liner  1014   b  is midpoint between the opposing pistons associated with the cylinder liner  1014   b , the cylinder liner  1014   b  is not currently undergoing scavenging. Accordingly, the exhaust gas passageways  1037  in the engine block  1013  are not yet configured in fluid communication with the exhaust gas passageways  1036  (see  FIG. 6 ) of the cylinder liner  1014   b.    
   Referring now to  FIG. 32 , the components of a reduction outdrive system  1094  will now be described. The reduction outdrive system  1094  translates the reciprocating and rotational motion of the crank-cam  1016  to rotational motion about a centerline of a power take-off shaft  1084 . The reduction outdrive system  1094  includes an outdrive reduction gear  1082  and an outdrive gear  1086 . The outdrive reduction gear  1082  further includes internal gear teeth  1090  disposed along the peripheral cylindrical wall of an outdrive gear receiving recess  1096 . The outdrive reduction gear  1082  is rigidly coupled to a power take-off drive flange  1080  by well-known means, such as fasteners. The power take-off shaft  1084  is perpendicularly and concentrically attached to the power take-off drive flange  1080 . The centerline of the power take-off shaft  1084  is collinear with the second axis of rotation  1078 . The outdrive gear  1086  has external gear teeth  1088  shaped and dimensioned to communicate with the internal gear teeth  1090  of the outdrive reduction gear  1082 . The outdrive gear  1086  has a crank web  1070  receiving recess  1092  shaped and dimensioned to receive the circular shaped crank web  1070 . The crank web  1070  is rigidly coupled to the receiving recess  1092  of the outdrive gear  1086  by means well known in the art, such as by fasteners. 
   In light of the above description of the components of the reduction outdrive system  1094 , the operation of the reduction outdrive system  1094  will now be described. Referring to  FIGS. 33-38 , a letter A is used as an arbitrarily selected reference point on the outdrive gear  1086  and a letter B is used as an arbitrarily selected reference point on the outdrive reduction gear  1082 . A reference letter C marks the center point of crank journal  1072   b , and thus the cylinder liner  1014   b  (not shown), and reference letter D marks the center point of the crank journal  1072   a  and thus the cylinder liner  1014   a  (not shown). 
   Referring now to  FIG. 33 , the outdrive gear  1086  is disposed within the outdrive reduction gear  1082 , so that the external gear teeth  1088  of the outdrive gear  1086  intermesh with the internal gear teeth  1090  of the outdrive reduction gear  1082 . As the outdrive reduction gear  1082  and the outdrive gear  1086  rotate clockwise while intermeshing, reference point D on the outdrive gear  1086  reciprocates along a horizontal reference line  1098 . The reference line  1098  represents the linear path of the cylinder liner  1014   b  (not shown) and is the same reference line depicted in  FIGS. 16-31 . Likewise, reference point C reciprocates along a vertical reference line  1100 . Vertical reference line  1100  represents the linear path of the cylinder liner  1014   a  (not shown) and is the same reference line depicted in  FIGS. 16-31 . As the outdrive reduction gear  1082  and outdrive gear  1086  rotate clockwise, reference point D moves to the right and reference point C moves upward, along their reference lines  1098  and  1100 , respectively. 
   Referring now to  FIG. 34 , the outdrive gear  1086  has rotated one-eighth of a turn clockwise while the outdrive reduction gear  1082  has rotated one-sixteenth of a turn clockwise from the configuration depicted in  FIG. 33 . As is apparent from reference to  FIG. 34 , reference points C and D still lie upon their respective reference lines  1100  and  1098 , thereby maintaining the linear path of travel of the centers of the crank journals and, thus, their attached cylinder liners. 
   Referring to  FIG. 35 , the outdrive gear  1086  has now rotated one-quarter of a turn clockwise, while the outdrive reduction gear  1082  has rotated one-eighth of a turn clockwise from the configuration depicted in  FIG. 33 . By referring to  FIG. 35 , it is apparent that reference point C has moved vertically upward along the linear reference line  1100 , while reference point D has moved horizontally to the right along the horizontal reference line  1098  from their respective positions depicted in  FIG. 34 . Reference point D is currently at its “zenith”; therefore, the respective cylinder liner is in an extended position, with the cylinder liner at a TDC and BDC position with reference to the substantially stationary opposing pistons associated with the cylinder liner. As the outdrive gear  1082  is rotated further clockwise, reference point D transitions from a rightward direction of travel to a leftward direction of travel along the reference line  1098 . 
   Referring now to  FIG. 36 , the outdrive gear  1086  has rotated one-half turn and the outdrive reduction gear  1082  has rotated one-quarter turn. Reference point C is now at its zenith; therefore, the corresponding cylinder liner is in an extended position with the cylinder liner at its TDC and BDC position with respect to the two substantially stationary opposing pistons associated with the cylinder liner. As the outdrive gear  1082  is rotated further clockwise, reference point C transitions from a upward direction of travel to a downward direction of travel along the reference line  1100 . 
   Referring now to  FIG. 37 , the outdrive gear  1086  has rotated three-quarters of a turn. The outdrive reduction gear  1082  has rotated three-eighths of a turn. Reference point C is now at the center of the reference path  1100 . This center position indicates that the cylinder liner associated with reference point C is now equidistant from the substantially stationary pistons associated with the cylinder liner. Correspondingly, reference point D is now at a zenith. Therefore, the cylinder liner associated with reference point D is at an extended position and thus, at a TDC and BDC position with regard to the substantially stationary opposing pistons associated with the cylinder liner. 
   Referring now to  FIG. 38 , the outdrive gear  1086  has rotated one full turn while the outdrive reduction gear  1082  has rotated one-half turn, as indicated by the relative positions of the reference points A and B. In one full rotation of the outdrive gear  1086 , each individual piston has gone through one complete thermodynamic cycle. Through the manipulation of diameters and the possible amount of gear teeth involved, different reduction ratios of engine RPM to power take-off shaft  1084  RPM are possible as should be apparent to one skilled in the art. In the illustrated embodiment depicted in  FIGS. 33-38 , the outdrive gear  1086  has 30 teeth and the outdrive reduction gear  1082  has 40 teeth. In one 360° rotation of the outdrive gear  1086 , the outdrive gear  1086  cams 60 teeth of the outdrive reduction gear  1082 . The outdrive reduction gear  1082  has 40 teeth; therefore, it rotates in the process the distance of 20 teeth, which results in a 180° rotation of the outdrive reduction gear  1082  and attached shaft. Thereby a ratio of 2:1 reduction in RPM is accomplished. 
   Often it is desirable to have a direct outdrive shaft that rotates at the same RPM as the engine or more specifically, at the crank-cam RPM. The direct outdrive shaft may be used to drive accessories, such as a distributor. Referring to  FIGS. 39-41 , a direct outdrive system  1102  formed in accordance with and suitable for use with the present invention is illustrated. The direct outdrive system  1102  includes a direct outdrive adapter  1104 , a direct outdrive  1106 , a direct outdrive shaft  1108 , and a gliding block  1110 . These components work in combination to convert the rotating and reciprocating motion of the crank-cam to a rotational movement in the direct outdrive output shaft  1108 . 
   The configuration of the direct outdrive adapter  1104  will now be discussed. The direct outdrive adapter  1104  is a disk-shaped member having inner (facing the engine) and outer (facing away from the engine) annular surfaces  1114  and  1116 , respectively. Formed adjacent to the inner annular surface  1114  is a crank web receiving recess  1118  where one of the crank webs  1070  (see  FIG. 14 ) is received and rigidly fastened there within. Perpendicularly and concentrically mounted relative to the outer annular surface  1116  is a drive shaft  1112 . The drive shaft  1112  is received within a bore  1120  located within the gliding block  1110 . 
   The configuration of the gliding block  1110  will now be discussed. The gliding block  1110  is generally a rectangular-shaped block structure having arcuate ends  1122  formed to match the outer circular circumference of the direct outdrive  1106 . The length and width of the gliding block  1110  is selected to match the length and width of a channel  1124  formed in the direct outdrive  1106 , thereby allowing the gliding block  1110  to be received within the channel  1124 . Preferably, a polished finish is applied to the contact surfaces of both the gliding block  1110  and the channel  1124  of the direct outdrive  1116  of which it rides within, to reduce friction and wear. 
   The direct outdrive  1106  is a disk-shaped member having inner (facing the engine) and outer (facing away from the engine) circular planar surfaces  1126  and  1128 , respectively. The channel  1124  for receiving the gliding block  1110  is formed on the inner planar surface  1126 . A direct drive output shaft  1108  is perpendicularly and concentrically mounted on the outer planar surface  1128 . 
   The operation of the direct outdrive system  1102  will now be described in reference to  FIGS. 42-45 . Referring now to  FIG. 42 , a planar end view of the direct outdrive system  1102  is shown, depicting the inner planar surface  1114  of the direct outdrive adapter  1104  with the crank-cam removed and the inner circular planar surface  1126  of the direct outdrive  1106 . The drive shaft  1112  of the adapter  1104  is shown in phantom. The gliding block  1110  is shown; however the majority of the gliding block  1110  is obscured by the adapter  1104 . The letter A is an arbitrarily selected reference point on the outer circumference of the direct outdrive  1106 , and the letter B is an arbitrarily selected reference point on the direct outdrive adapter  1104 . 
   Still referring to  FIG. 42 , the center of the direct outdrive adapter  1104  is indicated by reference numeral  1130 . The center of the direct outdrive  1106  is indicated by reference numeral  1132 . The direct outdrive adapter  1104  rotates about its center  1130 , while also revolving around the center  1132  of the direct outdrive  1106  along a circular orbit  1134 , the circular orbit  1134  having a radius equal to ¼ of the stroke length. 
     FIG. 43  shows the direct outdrive system  1102  rotated ¼ of a turn counterclockwise from that depicted in  FIG. 42 .  FIG. 44  shows the direct outdrive system  1102  rotated ½ of a turn counterclockwise from that depicted in  FIG. 42 .  FIG. 45  shows the direct outdrive system  1102  rotated ¾ of a turn counterclockwise from that depicted in  FIG. 42 . Inasmuch as the reference letters A and B remain radially aligned during the rotation of the direct outdrive adapter  1104  and direct outdrive  1106 , as shown in  FIGS. 42-45 , it should be apparent to one skilled in the art that both the adapter  1104  and the direct outdrive  1106  rotate at the same rate. Therefore, the direct outdrive output shaft  1108  (see  FIG. 41 ) may be used to drive components requiring rotary input rotating at engine RPM. 
   From examination of  FIGS. 42-45 , it appears that the sliding block  1110  does not move during operation. This would be true if the parts of the engine were constructed so as to have zero tolerances, i.e., if all parts were perfectly made exactly to specification. However, in the event the parts are constructed or wear so as to be within certain selected tolerances (i.e., plus or minus 10 thousands of an inch from a selected dimension), as is typically the case, the sliding block  1110  would undergo slight movements within the channel  1124 , thereby “absorbing” the tolerances of the parts, mitigating vibration and reducing the potential of the parts&#39; binding. 
   Like all internal combustion engines, the illustrated reciprocating internal combustion engine  1010  produces large amounts of heat during operation, most of it as a result of the combustion process, additional heat being generated by the compression of the gases within the cylinder liners and the friction between the moving parts of the engine  1010 . Temperatures within the engine  1010  are kept under control by a cooling system that circulates coolant through passages in the engine block and around critical parts to remove excess heat and to equalize stresses produced by heating. Inasmuch as the design and components of internal combustion engine cooling systems are well known in the art, the cooling passages in the engine and cooling system components are not shown for the purpose of clarity. 
   Referring to  FIGS. 46-53 , one embodiment of a power transfer assembly  2000  formed in accordance with the present invention and suitable for use in transferring power generated in an internal combustion engine externally of the engine for use is provided. The illustrated embodiment, although illustrated and described for use with the combustion engine depicted in and described with reference to  FIGS. 1-45 , may also be used in all types of combustion engines, including those with stationary pistons, such as disclosed in U.S. Pat. Nos. 6,598,567 and 6,032,622, the disclosures of which are hereby expressly incorporated by reference, and more conventional engine designs having stationary cylinders and moving pistons, as should be apparent to those skilled in the art. 
   Turning to  FIG. 46 , generally stated, the power transfer assembly  2000  converts the linear motion of the cylinders (or pistons in more conventionally designed combustion engines) to rotary motion, and permits the transfer of the power generated by the combustion of fuel in the engine externally of the engine for use, such as providing power to drive the wheels of a vehicle. 
   The power transfer assembly  2000  of  FIGS. 46-53  includes two subassemblies; a crankshaft assembly  2002  and an outdrive assembly  2004 . The crankshaft assembly  2002  is substantially similar to the crank-cam  1016  depicted in  FIGS. 12-15  with a couple of exceptions which will be described in more detail below. The outdrive assembly  2004  is substantially similar to the reduction outdrive system  1094  and direct outdrive system  1102  depicted in  FIGS. 32-45 , with the exception that the outdrive assembly  2004  includes means for improved balancing of vibrations in the power transfer assembly  2000  and provides the ability to provide both reduced and direct power output on one side of the engine. Inasmuch as the components and operation of the power transfer assembly  2000  of  FIGS. 46-53  are substantially similar to the corresponding components of the embodiment described in relation to  FIGS. 1-45 , for the sake of brevity, this detailed description will focus only upon those aspects of structure and operation which depart from the embodiment described above. 
   Still referring to  FIG. 46 , and as stated above, the power transfer assembly  2000  includes a crankshaft assembly  2002 . The crankshaft assembly  2002  includes a crankshaft  2006 , which in the illustrated embodiment, is in the form of a “crank-cam,” labeled as such since the crankshaft  2006  includes a pair of cams  2008  disposed on the crankshaft for valve actuating as discussed above. The crankshaft  2006  of this embodiment is substantially similar to the crank-cam  1016  of  FIGS. 1-45  with the exception that the crankshaft  2006  includes a power transfer device, such as a power transfer gear  2010 , and a rotation connection assembly for permitting a crankshaft counterweight to be rotatably coupled to the crankshaft, the rotation connection assembly of the illustrated embodiment being in the form of a pair of stub shafts  2012  which rotatingly receive a pair of crankshaft counterweights. The gear  2010  and stub shafts  2012  are concentrically located relative to a rotational axis  2014  of the crankshaft  2006 . One of the stub shafts  2012  extends outward from the right most crank web  2018  and outward of the gear  2010 . The other stub shaft  2012  extends outward from the left most crank web  2016 . The stub shafts  2012  and the gear  2010  assist in transfer of power from the crankshaft  2006  to the outdrive assembly  2004  as will be described in further detail below. 
   Still referring to  FIG. 46 , the outdrive assembly  2004  includes two subassemblies; a direct outdrive assembly  2020  and a direct and reduction outdrive assembly  2022 . Focusing on the direct outdrive assembly  2020 , the direct outdrive assembly  2020  converts the rotating and reciprocating motion of the crankshaft  2006  to rotary motion in a direct outdrive output shaft  2024 . The direct outdrive output shaft  2024  rotates at the same RPM as the engine or more specifically, at the same RPM that the crankshaft  2006  orbits about an axis of rotation  2038  of the outdrive assembly  2004 . The direct outdrive output shaft  2024  may be used to drive items located externally of the engine, such as accessories, one suitable example being a distributor. The direct outdrive assembly  2020  includes a direct outdrive  2026 , the direct outdrive shaft  2024 , and a balancing gliding block  2028 . 
   Turning to  FIGS. 46-48 , the configuration of the balancing gliding block  2028  will now be discussed. The balancing gliding block  2028  includes a follower  2029  that is generally a rectangular-shaped block structure having arcuate ends  2030  formed to match the outer circular circumference of the direct outdrive  2026 . The width of the follower  2029  is selected to match the width of a guide or channel  2032  formed in the direct outdrive  2026 , thereby allowing the follower  2029  to be slidably received within the channel  2032 . Preferably, a polished finish is applied to the contact surfaces where the follower  2029  and the channel  2032  contact each other to reduce friction and wear. 
   The follower  2029  in combination with the channel  2032  form a slide assembly, which is part of an interface assembly  2031 . The interface assembly  2031  is used for non-rigidly interfacing the crankshaft  2006  with a pair of output shafts  2024  and  2046  to facilitate the direct transfer of torque between the crankshaft  2006  and the output shafts  2024  and  2046  while still permitting the crankshaft  2006  to move relative to the output shafts  2024  and  2046  to absorb any misalignment of the crankshaft  2006 . Moreover, the crankshaft  2006  is able to rotate relative to the direct outdrive  2026  or  2066  about at least one axis (i.e. axis  2014  going through the stub shafts  2012 ) and move freely relative to the outdrive  2026  or  2066  in at least one direction, one suitable direction being linearly outward from a center axis of the direct outdrive  2026  such as indicated by axis  2027  in  FIG. 47 . Of note, axis  2027  is oriented to always pass through the center axis  2014  of the crankshaft  2006  and the center axis  2038  of the outdrive assembly  2004 . This arrangement permits a torque to be directly transferred from the crankshaft  2006  to the output shafts  2024  and  2046  without a centrifugal force present in the crankshaft  2006  (from its orbital movement) and/or counterbalance weight  2036  being transferred from the crankshaft  2006  and/or counterbalance weight  2036  to the output shafts  2024  and  2066  since the crankshaft  2006  and counterbalance weight  2036  are free to slide radially outward relative to the direct outdrive  2026  along axis  2027 . 
   The balancing gliding block  2028  further includes a bore  2034  disposed perpendicularly through the follower  2029 . The bore  2034  is sized and positioned to rotatingly receive one of the stub shafts  2012  of the crankshaft  2006 . The balancing gliding block  2028  still further includes a counterbalance weight  2036 . The counterbalance weight  2036  is sized and positioned to counterbalance the crankshaft  2006  as the crankshaft  2006  orbits about the axis of rotation  2038  of the outdrive assembly  2004  as will be described in more detail below. 
   The direct outdrive  2026  is a disk-shaped member having inner (facing the engine) and outer (facing away from the engine) circular planar surfaces  2040  and  2042 , respectively. The channel  2032  for receiving the follower  2029  of the balancing gliding block  2028  is formed on the inner planar surface  2040 . A direct drive output shaft  2024  is perpendicularly and concentrically mounted on the outer planar surface  2042 . 
   Turning now to  FIGS. 49-51 , the components of the direct and reduction outdrive assembly  2022  will now be described. The direct and reduction outdrive assembly  2022  converts the rotating and reciprocating motion of the crankshaft  2006  to rotary motion in a pair of concentrically disposed output shafts  2044  and  2046 . One of the output shafts, which will be referenced as a reduction output shaft  2044 , rotates at a reduced RPM relative to the engine or more specifically, relative to the RPM that the crankshaft  2006  orbits about the axis of rotation  2038  of the outdrive assembly  2004 . The reduction output shaft  2044  may be used to drive items external of the engine, such as a drive wheel of a vehicle. The direct and reduction outdrive assembly  2022  also includes a direct output shaft  2046  concentrically located within the reduction output shaft  2044 . The direct output shaft  2046  rotates at the same RPM as the engine, or more specifically, at the same RPM that the crankshaft  2006  orbits about the axis of rotation  2038  of the outdrive assembly  2004 . Thus, the direct and reduction outdrive assembly  2022  provides both a reduced RPM output shaft  2044  and a direct output shaft  2046  on one side of the engine. Of note, the output shafts  2044  and  2046  rotate in opposite directions relative to one another. 
   The direct and reduction outdrive assembly  2022  includes an outdrive reduction gear  2048  and an outdrive gear  2050 . The outdrive reduction gear  2048  further includes internal gear teeth  2052  disposed along the inner peripheral cylindrical wall of an outdrive hub  2054 . The outdrive hub  2054  is rigidly coupled or integrally formed with the reduction output shaft  2044 . The reduction output shaft  2044  is perpendicularly and concentrically attached to the outdrive hub  2054  and reduction gear  2048 . The centerline of the outdrive hub  2054  is collinear with the axis of rotation  2038  of the outdrive assembly  2004 . The outdrive hub  2054  includes a cavity  2068  shaped and sized to permit a balancing gliding block  2062  and direct outdrive  2066  to rotate freely therein. 
   The outdrive gear  2050  has external gear teeth  2056  shaped and dimensioned to communicate with the internal gear teeth  2052  of the outdrive reduction gear  2048 . The outdrive gear  2050  has a crank web  2018  receiving recess  2058  shaped and dimensioned to receive the circular shaped crank web  2018 . The outdrive gear  2050  further includes internal gear teeth  2060  shaped and dimensioned to communicate with the external gear teeth of the gear  2010  disposed on the crankshaft  2006  such that any rotation of the crankshaft  2006  about its axis of rotation  2014  will result in a corresponding rotation of the outdrive gear  2050 . The size of and the number of teeth of the outdrive reduction gear  2048  and the outdrive gear  2050  may be selected to determine the amount of reduction desired. In the illustrated embodiment, the size and number of teeth are selected for a 2:1 reduction ratio. Although a specific reduction gear ratio is illustrated and described, it should be apparent to those skilled in the art that other reduction gear ratios, either higher or lower than described above, are suitable for use with and within the spirit and scope of the present invention. Of note, the backlash of gears  2048  and  2050  is selected to absorb any misalignment of crankshaft  2006  in a manner well known in the art. 
   The direct and reduction outdrive assembly  2022  further includes a balancing gliding block  2062  formed substantially as described above with reference to the balancing gliding block  2028  of the direct outdrive  2020  (see  FIGS. 46-48 ), with the exception of the shape of the counterbalance weight  2064 , which has been made thinner and wider to be more compact. 
   The direct and reduction outdrive assembly  2022  further includes a direct outdrive  2066  which is substantially identical to the direct outdrive  2026  of  FIGS. 46-48 . The direct outdrive  2066  interfaces with the balancing gliding block  2062  and crankshaft  2006  in the same manner that the direct outdrive  2026  and balancing gliding block  2028  of  FIGS. 46-48  do. 
   In light of the above description of the components of the power transfer assembly  2000  and turning to  FIGS. 46 and 52A , the operation of the power transfer assembly  2000  will now be described.  FIG. 52A  is a cross-sectional view of the power transfer assembly  2000  of  FIG. 46 , the cross-sectional cut taken substantially through Section  52 A- 52 A of  FIG. 46 .  FIG. 52A  shows the balancing glide block  2028 , direct outdrive  2026 , the crankshaft rotational axis  2014  about which the crankshaft  2006  rotates, the orbital path  2070  of the crankshaft rotational axis  2014  about the axis of rotation  2038  of the outdrive assembly  2004 , and a first crank journal  2072  (shown in cross-hatching) and a second crank journal  2074  (shown in phantom). 
   During operation, the first crank journal  2072  reciprocates (first towards the top of the page and subsequently towards the bottom of the page) along a vertical axis  2076  moving a cylinder (not shown) there along while the second crank journal  2074  reciprocates (first to the right of the page and subsequently towards the left of the page) along a horizontal axis  2078  moving a second cylinder (not shown) there along. The direct outdrive  2026  and balancing gliding block  2028  rotate in a counterclockwise direction at the same rate as each other. The crankshaft  2006  rotates about the crankshaft rotational axis  2014  in a clockwise direction while the crankshaft rotational axis  2014  (and thus the crankshaft  2006 ) orbits about the axis of rotation  2038  of the outdrive assembly  2004  in a counterclockwise manner along the orbital path  2070 . 
   One of the crankshaft  2006  stub shafts  2012  is rotatingly interfaced with the crankshaft counter weight such that the stub shaft  2012  is permitted to freely rotate within the balancing glide block  2028 . This permits the crankshaft  2006  to rotate about the crankshaft rotation axis  2014  during operation. Further, the channel  2032  is oriented in a linear path located radial outward from the axis of rotation  2038  of the outdrive assembly  2004 . This permits the crankshaft  2006  and counterbalance weight  3036  to move radially outward from the axis of rotation  2038  of the outdrive assembly  2004  during operation. This permits any misalignments in the crankshaft  2006  orientation or associated components to be “absorbed” such that binding of the engine is impeded. Of note, the crankshaft  2006  and counterbalance weight  2036  are interfaced with the outdrive assembly  2004  such that they are permitted to move “freely” in a selected direction, the selected direction being radially outward in the illustrated embodiment. The term “freely” meaning, for the purposes of this detailed description, that the crankshaft and/or crankshaft counterweight are not biased or restricted from moving in the selected direction. Thus, it can be seen from the above that the crankshaft  2006  and counterweight  2036  are non-rigidly interfaced with the outdrive assembly  2004  such that the crankshaft  2006  is permitted to freely rotate relative to the direct outdrive  2026  in at least one direction and freely move radially relative to the direct outdrive  2026  during operation in at least one direction. 
   During operation, the counterbalance weight  2036  is always disposed directly opposite the crankshaft rotation axis  2014  to counterbalance the centrifugal forces produced by the mass of the crankshaft  2006  being offset from the axis of rotation  2038  of the outdrive assembly  2004  as the crankshaft  2006  orbits along the orbital path  2070 . Referring to  FIG. 52B , the weight of the counterbalance weight  2036  and the location of the center of mass  2015  of the counterbalance weight  2036  is oriented relative to the crankshaft rotational axis  2014  such that when the crankshaft is rotated the centrifugal force associated with the counterbalance weight  2036  balances the orbital centrifugal force of the crankshaft  2006 . More specifically, during operation, the center of mass  2015  of the counterbalance weight  2036  is preferably oriented along an axis  2027  oriented to pass through both the center axis  2014  about which the crankshaft  2006  rotates and through the axis  2038  about which the crankshaft  2006  orbits. In this configuration, the mass of the counterbalance weight  2036  is able to balance the centrifugal forces generate by the orbiting of the crankshaft  2006  about axis  2038 . 
   The interrelationship of the parts of the power transfer assembly  2000 , and more specifically of the direct outdrive assembly  2020 , is best understood by examination of  FIGS. 52A-52H .  FIGS. 52A-52H  show the direct outdrive assembly  2020  as the direct outdrive  2026  rotates through one cycle. More specifically,  FIG. 52B  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52A .  FIG. 52C  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52B .  FIG. 52D  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52C .  FIG. 52E  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52D .  FIG. 52F  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52E .  FIG. 52G  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52F .  FIG. 52H  depicts the direct outdrive assembly  2020  as the direct outdrive  2026  rotates 45 degrees counterclockwise from the position of the direct outdrive  2026  shown in  FIG. 52G . Once the direct outdrive  2026  has been rotated 45 degrees counterclockwise from the position shown in  FIG. 52H , the direct outdrive  2026  has completed one full revolution, returning to the position shown in  FIG. 52A . 
   Referring to  FIGS. 52A-52H , during one cycle of the power transfer assembly  2000 , it can be seen, among other things, that the following occurs: 1) the direct outdrive  2026  rotates once in the counterclockwise direction; 2) the crankshaft  2006  rotates once in the clockwise direction about the crankshaft rotational axis  2014 ; 3) the crankshaft  2006  orbits once in the counterclockwise direction about the axis of rotation  2038  of the outdrive assembly  2004  about the orbital path  2070 ; 4) the crank journals  2072  and  2074  move along linear paths  2076  and  2078  perpendicular to one another; 5) the stroke length of the cylinders is defined by the distance moved by the “vertical” crank journal  2072  when moved from the bottom of the page to the top of the page, or by the distance moved by the “horizontal” crank journal  2074  when moved from the right most position on the page to the crank journal&#39;s  2074  left most position; 6) the counterbalance weight  2036  is rotatingly coupled to the crankshaft  2006 ; 7) the counterbalance weight  2036  is slideable along a path (defined by the channel  2032  in the direct outdrive  2026 ) oriented radially relative to an axis of rotation  2038  of the outdrive assembly  2004 ; 8) the orbital direction of rotation of the crankshaft  2006  is the same direction of rotation of the direct outdrive  2026 ; 9) the crankshaft  2006  is not rigidly attached nor attached in a biased manner to the direct outdrive  2026 ; 10) the centerline of the crankshaft  2006  is located a ¼ of a stroke length outward from the centerline of the direct outdrive  2026 ; and 11) the Revolutions Per Minute (RPM) of the crankshaft  2006  about its own axis is the same as the RPM of the orbital movement of the crankshaft  2006 ; 12) the counterbalance weight  2036  rotates at an RPM equal to both the RPM of the crankshaft  2006  about its own centerline and at the same RPM as the orbital movement of the crankshaft  2006  about the centerline of the direct outdrive  2026 ; 13) the counterbalance weight  2036  rotates in the same rotary direction as the orbital movement of the crankshaft  2006  about the centerline of the direct outdrive  2026 ; and 14) the counterbalance weight  2036  rotates in the opposite rotary direction relative to the rotary movement of the crankshaft  2006  about its own centerline. 
     FIG. 53A  is a cross-sectional view of the power transfer assembly  2000  of  FIG. 46 , the cross-sectional cut taken substantially through Section  53 A- 53 A of  FIG. 46 .  FIG. 53A  shows the balancing glide block  2062 , crank web  2018 , outdrive gear  2050 , direct outdrive  2066 , the crankshaft rotational axis  2014  about which the crankshaft  2006  rotates, the orbital path  2070  of the crankshaft rotational axis  2014  about the axis of rotation  2038  of the outdrive assembly  2004 , and a first crank journal  2072  (shown in phantom) and a second crank journal  2074  (shown in cross-hatching). 
   During operation, the first crank journal  2072  reciprocates (first towards the top of the page and subsequently towards the bottom of the page) along a vertical axis  2076  moving a cylinder (not shown) there along while the second crank journal  2074  reciprocates (first to the left of the page and subsequently towards the right of the page) along a horizontal axis  2078  moving a second cylinder (not shown) there along. The direct outdrive  2066  (see  FIG. 46 ) and balancing gliding block  2062  rotate in a clockwise direction at the same rate as each other. The crankshaft  2006  rotates about the crankshaft rotational axis  2014  in a counterclockwise direction while the crankshaft rotational axis  2014  (and thus the crankshaft  2006 ) orbits about the axis of rotation  2038  of the outdrive assembly  2004  in a clockwise manner along the orbital path  2070 . During operation, the counterbalance weight  2064  is always disposed directly opposite the crankshaft rotation axis  2014  to counterbalance the centrifugal forces produced by the mass of the crankshaft  2006  being offset from the axis of rotation  2038  of the outdrive assembly  2004  as the crankshaft  2006  orbits about the orbital path  2070  in the same manner as described above for the outdrive assembly  2004 . 
   The interrelationship of the parts of the power transfer assembly  2000  is best understood by examination of  FIGS. 53A-53H .  FIGS. 53A-53H  show the power transfer assembly  2000  as the direct outdrive  2066  of the direct and reduction outdrive assembly  2022  rotates through one cycle. More specifically,  FIG. 53B  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53A  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  on the crankshaft  2006  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise. As should be apparent to those skilled in the art, the relative movement of the gears  2050  and  2048  relative to one another, i.e., the reduction ratio, may be adjusted through adjusting the number of teeth on the gears  2050  and  2048 . The relative movement of the outdrive gear  2050  relative to the outdrive reduction gear  2048  is best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048  respectively in  FIGS. 53A-53H . 
     FIG. 53C  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53B  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . 
     FIG. 53D  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53C  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . 
     FIG. 53E  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53D  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . 
     FIG. 53F  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53E  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive gear  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . 
     FIG. 53G  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53F  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . 
     FIG. 53H  depicts the power transfer assembly  2000  as the direct outdrive  2066  rotates 45 degrees clockwise from the position of the direct outdrive  2066  shown in  FIG. 53G  as best shown by examining the change in position of the reference point “C” marked on the direct outdrive  2066 . While the direct outdrive  2066  rotates 45 degrees to the clockwise, the outdrive gear  2050  rotates 45 degrees counterclockwise about the crankshaft rotational axis  2014 . The 45 degree rotation of the outdrive gear  2050  causes the outdrive reduction gear  2048  to rotate 22.5 degrees counterclockwise as best shown by examining the change in position of reference points “A” and “B” marked on the gears  2050  and  2048 . Once the direct outdrive  2066  has been rotated 45 degrees clockwise from the position shown in  FIG. 53H , the direct outdrive  2066  has completed one full revolution in the clockwise direction, returning to the position shown in  FIG. 53A . The outdrive reduction gear  2048  will have rotated one half revolution in the counterclockwise direction. 
   Referring to  FIGS. 53A-53H , during one cycle of the power transfer assembly  2000 , it can be seen, among other things, that the following occurs: 1) the direct outdrive  2026  rotates once in the clockwise direction; 2) the crankshaft  2006  rotates once in the counterclockwise direction about the crankshaft rotational axis  2014 ; 3) the crankshaft  2006  orbits once in the clockwise direction about the axis of rotation  2038  of the outdrive assembly  2004  about the orbital path  2070 ; 4) the crank journals  2072  and  2074  move along linear paths  2076  and  2078  perpendicular to one another; 5) the stroke length of the cylinders is defined by the distance moved by the “vertical” crank journal  2072  when moved from the bottom of the page to the top of the page, or by the distance moved by the “horizontal” crank journal  2074  when moved from the left most position on the page to the crank journal&#39;s  2074  right most position; 6) the counterbalance weight  2064  is rotatingly coupled to the crankshaft  2006 ; and 7) the counterbalance weight  2064  is slideable along a path oriented radially relative to an axis of rotation  2038  of the outdrive assembly  2004 ; 8) the orbital direction of rotation of the crankshaft  2006  is the same direction of rotation of the direct outdrive  2066 ; 9) the crankshaft  2006  is not rigidly attached nor attached in a biased manner to the direct outdrive  2066 ; 10) the centerline of the crankshaft  2006  is located at ¼ of a stroke length outward from the centerline of the direct outdrive  2066 ; 11) the Revolutions Per Minute (RPM) of the crankshaft  2006  about its own axis is the same as the RPM of the orbital movement of the crankshaft  2006 ; 12) the counterbalance weight  2064  rotates at an RPM equal to both the RPM of the crankshaft  2006  about its own centerline and at the same RPM as the orbital movement of the crankshaft  2006  about the centerline of the direct outdrive  2066 ; 13) the counterbalance weight  2064  rotates in the same rotary direction as the orbital movement of the crankshaft  2006  about the centerline of the direct outdrive  2066 ; 14) the counterbalance weight  2064  rotates in the opposite rotary direction relative to the rotary movement of the crankshaft  2006  about its own centerline; 15) the output shafts  2038  rotate opposite one another, with the reduced output shaft  2038  rotating opposite the orbital movement of the crankshaft  2006  and in the same direction as the rotation of the crankshaft  2006  about its own centerline; and 16) the rotary and orbital movement of the crankshaft  2006  combine to drive the reduced output shaft  2038  at a reduced RPM relative to the direct output shaft  2046 . 
   Referring to  FIG. 54 , an alternate embodiment of a power transfer assembly  3000  formed in accordance with the present invention is shown. The power transfer assembly  3000  of  FIG. 54  is substantially similar in structure and operation to the power transfer assembly  2000  of  FIGS. 46-53 . Therefore, for the sake of brevity, this detailed description will focus only upon the aspects of the alternate embodiment which depart from the embodiment illustrated and described above. 
   That said, the power transfer assembly  3000  departs from that described above in that the diameter of the direct outdrives  2026  and  2066  of the previously described power transfer assembly  2000 , best shown in  FIG. 46 , have been reduced. The reduction in diameter of the direct outdrives  3026  and  3066  of the power transfer assembly  3000  of  FIG. 54  permits the counterbalance weights  3036  and  3064  to be increased in size by extending radially outward of the direct outdrives  3026  and  3066  so as to overlap the outer edges  3004  of the direct outdrives  3026  and  3066 . 
   Referring to  FIGS. 55 and 56 , an alternate embodiment of a direct outdrive assembly  4020  formed in accordance with the present invention is shown. The direct outdrive assembly  4020  is substantially similar in structure and operation to the direct outdrive assembly  2020  of  FIGS. 47 and 48 . Therefore, for the sake of brevity, this detailed description will focus only upon the aspects of the alternate embodiment which depart from the embodiment illustrated and described above. 
   That said, the direct outdrive assembly  4020  departs from that described above in that the follower  4029  of the interface assembly  4031  is now located on the direct outdrive  4026  and the channel  4032  for slidingly receiving the follower  4029  is now located on the balancing gliding block  4028 . Thus, the positions of the follower  4029  and the channel  4032  have been swapped. It should be apparent to those skilled in the art that the swapping of the locations of the follower and channel is also suitable for the direct outdrive components of the direct and reduction outdrive assembly  2022 , best shown in  FIGS. 46 ,  50 , and  51 . 
   Referring to  FIG. 46 , although the above described embodiment is illustrated and described as having a pair of stub shafts  2012  coupled to the crankshaft  2006 , it should be apparent to those skilled in the art that the stubshafts  2012  are optional components and an engine formed in accordance with the present invention may operate without the use of one or more of the stubshafts  2012 . For instance, in the embodiment depicted in  FIG. 32 , the crankshaft does not utilize stubshafts. Further, referring to  FIG. 41 , if a stubshaft  1112  is desired, an adapter  1102  may be coupled to the crankshaft to provide a stubshaft. Thus, it should be apparent to those skilled in the art that a crankshaft having one stubshaft, or no stubshafts, or one that uses adapters to provide one or more stubshafts, are also within the spirit and scope of the present invention. 
   While the preferred embodiment of the invention has been illustrated and described, it will be appreciated that various changes can be made therein without departing from the spirit and scope of the invention.