Patent Publication Number: US-2018051626-A1

Title: Variable geometry power transfer for fluid flow machines

Description:
This application is a divisional of U.S. patent application Ser. No. 14/281,352 filed May 19, 2014, which claims priority to U.S. Provisional Application No. 61/825,362, filed May 20, 2013 and entitled Efficient Variable Geometry Power Transfer For Combustion Engines; and U.S. Provisional Application No. 61/897,011, filed Oct. 29, 2013 and entitled Efficient Variable Geometry Power Transfer For Combustion Engines, the entirety of all applications are incorporated by reference herein. 
    
    
     BACKGROUND OF THE INVENTION 
     A combustion engine is an engine in which the combustion of fuel and air occurs within a combustion chamber. The combustion process burns the fuel and air mixture to create a gas at high temperature. The high temperature gas creates high pressure that is then used to apply force to a piston to perform work. Because the combustion process generates a gas, the ideal gas law can be used to determine the relationship between temperature, pressure, and volume of the gas. 
     The ideal gas law is PV=nRT, where 
     P=pressure 
     V=volume of the gas 
     T=temperature 
     n=number of moles of gas 
     R=ideal gas constant 
     Given a constant quantity of gas, the pressure of the gas is directly related to its temperature and inversely related to its volume. 
     Most combustion engines in use today use a crank slider mechanism (CSM) to transfer the power from the linear motion of the piston to circular motion. The CSM includes a piston connected to a crankshaft (crank) by a connecting rod, as is shown in  FIG. 1 . In  FIG. 1 , F W  is the force perpendicular to the crank and generates torque T on the crank. F P  is the force on the piston caused by the combustion process. The relationship between T, F W  and F P  is shown as follows: 
         T=F   W *(stroke/2) 
     F W  is related to F P  by the following equation 
         F   W   =F   P *sin(180−β− a )
 
       FIG. 2  is a graph showing the force F W  as a percentage of force F P  as the crank rotates from angle θ at Top Dead Center (TDC) to 180 degrees at Bottom Dead Center (BDC) for a crank slider with a stroke of 4 inches and a connecting rod 6 inches in length. It can be seen from this graph that F W  is equal to 0 at TDC and increases until F W  is equal to F P  at 65 degrees of crank motion and then decreases until F W  is again 0 at 180 degrees. 
     According to the ideal gas law, the force on the piston (Fp) varies with the inverse of the volume of the gas.  FIG. 3  (Prior Art) is a graph showing the relationship between gas pressure and the crank angle for a typical combustion engine. In this chart, negative degrees are before TDC and positive degrees are after TDC with 0 being TDC. The pressure rises before TDC as the fuel mixture is being compressed into a smaller volume. In addition, for real systems, the fuel mixture takes a finite amount of time to burn requiring the fuel mixture be ignited before TDC. This can be seen in the graph as a change of the slope in the curve as the fuel mixture is ignited at −20 degrees before TDC.  FIG. 3  (Prior Art) shows that the pressure peak occurs 5 degrees past TDC but this peak can be moved by igniting the fuel earlier or later in the combustion process. The slope of the pressure decrease after peak is driven by the amount of additional volume in the combustion chamber as the crank rotates. 
       FIG. 4  (Prior Art) is a graph showing two curves of pressure versus crank angle. The first curve (solid line) has the pressure peak at 5 degrees after TDC and the second curve (dashed line) has the pressure peak at 20 degrees after TDC. The graph shows that the crank angle at which the power peak occurs can be changed but such comes at a price, as the volume of the gas increases the later the fuel mixture is ignited, resulting in a lower overall peak value. 
       FIG. 5  (Prior Art) is a graph comparing available pressure from the combustion process to the pressure that is converted to do work. The “Pressure Available” curve is derived from the ideal gas law with temperature constant and  100  being the force available at TDC. The pressure available drops as the crank rotates and the volume of the combustion chamber expands. The “CSM” curve is derived from multiplying the CSM percentage of force converted to work times the force available.  FIG. 5  shows that close to TDC, there is a great deal of pressure but very little of it is converted to work, which is shown as the large gap between the two curves at the lower crank angles. This comparison shows that at the pressure peak, 0% of the force is used to perform work and by the time 100% of the pressure is converted into work, the pressure is 25% of its peak. Because of this, a conventional crank slider mechanism only converts approximately one half of the available pressure into work. 
     Engines not having a conventional crank slider mechanism have been proposed in U.S. Pat. No. 6,684,828 to Ushijima; U.S. Pat. No. 7,213,563 to Yaguchi; U.S. Pat. No. 7,992,529 to Kobayashi; U.S. Pat. No. 8,011,343 to Kobayashi; U.S. Pat. No. 8,100,098 to Takahashi; U.S. Pat. No. 8,161,922 to Watanabe; U.S. Pat. No. 8,171,899 to Watanabe; U.S. Pat. No. 8,281,764 to Gurler and U.S. Pat. No. 8,327,819 to Voegeli. 
     BRIEF SUMMARY OF THE INVENTION 
     A fluid flow machine includes a casing including a cylinder and a crankshaft support. A piston is slidably disposed in the cylinder for reciprocating along an axis of the cylinder. A crankshaft includes a main bearing journal rotationally supported in the crankshaft support, a crankpin radially offset from an axis of the main bearing journal and a crank web connecting the main bearing journal and the crankpin. A multi-linkage connecting rod mechanism is connected between the piston and crankpin and includes a connecting rod, a first hinge link and a crankpin link pivotally connected to each other. A force transfer mechanism connects the multi-linkage connecting rod mechanism to the casing for transferring a vertical piston force into a horizontal crankpin force. 
     The various features of novelty which characterize the invention are pointed out with more particularity in the claims annexed to and forming a part of this disclosure. For a better understanding of the invention, its operating advantages and specific objects attained by its use, reference should be made to the accompanying drawings and descriptive matter in which there are illustrated and described preferred embodiments of the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Further objects and advantages of this invention will become more apparent and more readily appreciated from the following detailed description of the present invention, taken in conjunction with the accompanying drawings, of which: 
         FIG. 1  (Prior Art) shows a force diagram for a conventional piston combustion engine; 
         FIG. 2  (Prior Art) is a graph showing percent of piston force to work by crank angle; 
         FIG. 3  (Prior Art) is a graph showing the relationship between gas pressure and crank angle for a typical combustion engine; 
         FIG. 4  (Prior Art) is a graph showing two curves of pressure versus crank angle; 
         FIG. 5  (Prior Art) is a graph comparing available pressure from the combustion process to the pressure that is converted to do work; 
         FIG. 6  shows a force diagram of an embodiment of the machine of the present invention; 
         FIG. 7  shows a force diagram for a non-circular gear force transfer mechanism; 
         FIG. 8  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 9  is a schematic detail view of a multi-linkage connecting rod mechanism of the embodiment of  FIG. 8 ; 
         FIG. 10A  shows a first position in a sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 10B  shows a second position in the sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 10C  shows a third position in the sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 10D  shows a fourth position in the sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 10E  shows a fifth position in the sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 10F  shows a sixth position in the sequence of six positions for the engine of  FIG. 8  as the engine rotates through a revolution; 
         FIG. 11  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 12  is a schematic detail view of a multi-linkage connecting rod mechanism of the embodiment of  FIG. 11 ; 
         FIG. 13  is a schematic detail view of a force transfer mechanism of the embodiment of  FIG. 11 ; 
         FIG. 14A  shows a first position in a sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 14B  shows a second position in the sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 14C  shows a third position in the sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 14D  shows a fourth position in the sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 14E  shows a fifth position in the sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 14F  shows a sixth position in the sequence of six positions for the engine of  FIG. 11  as the engine rotates through a revolution; 
         FIG. 15A  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 15B  is a schematic sectional view of the embodiment of  FIG. 15A  taken along section line A-A; 
         FIG. 16  is a perspective view of a movable mount of the embodiment of  FIG. 15A ; 
         FIG. 17  is a schematic sectional view of the embodiment of  FIG. 15A ; 
         FIG. 18  is a schematic detail sectional view of the embodiment of  FIG. 15A ; 
         FIG. 19  is a perspective view of a cam follower of the embodiment of  FIG. 15A ; 
         FIG. 20  is a perspective view of a locking pin of the embodiment of  FIG. 15A ; 
         FIG. 21  is a schematic detail sectional view of the embodiment of  FIG. 15A  in a different position; 
         FIG. 22A  shows a first position in a sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 22B  shows a second position in the sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 22C  shows a third position in the sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 22D  shows a fourth position in the sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 22E  shows a fifth position in the sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 22F  shows a sixth position in the sequence of six positions for the engine of  FIG. 15A  as the engine rotates through a revolution; 
         FIG. 23A  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 23B  is a schematic sectional view of the embodiment of  FIG. 23A  taken along section line A-A; 
         FIG. 24  is a perspective view of a multi-linkage connecting rod mechanism of the embodiment of  FIG. 23A ; 
         FIG. 25A  shows a first position in a sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 25B  shows a second position in the sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 25C  shows a third position in the sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 25D  shows a fourth position in the sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 25E  shows a fifth position in the sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 25F  shows a sixth position in the sequence of six positions for the engine of  FIG. 23A  as the engine rotates through a revolution; 
         FIG. 26A  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 26B  is a schematic sectional view of the embodiment of  FIG. 26A  taken along section line A-A; 
         FIG. 27  is a schematic detail sectional view of the embodiment of  FIG. 26A ; 
         FIG. 28  is a perspective view of a movable mount of the embodiment of  FIG. 26A ; 
         FIG. 29  shows side and edge views of the movable mount of  FIG. 28 ; 
         FIG. 30  is a perspective view of a second side of the moving pivot point housing of the embodiment of  FIG. 26A ; 
         FIG. 31  is a perspective view of a first side of the moving pivot point housing of  FIG. 30 ; 
         FIG. 32  is a perspective view of the locking pin of the embodiment of  FIG. 26A ; 
         FIG. 33  is a perspective view of the second cam follower of the embodiment of  FIG. 26A ; 
         FIG. 34  is a schematic detail sectional view of the embodiment of  FIG. 26A ; 
         FIG. 35A  shows a first position in a sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 35B  shows a second position in the sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 35C  shows a third position in the sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 35D  shows a fourth position in the sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 35E  shows a fifth position in the sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 35F  shows a sixth position in the sequence of six positions for the engine of  FIG. 26A  as the engine rotates through a revolution; 
         FIG. 36  is a schematic sectional view of an embodiment of the present invention; 
         FIG. 37  is a schematic sectional perspective view of the embodiment of  FIG. 36  from a reverse side; 
         FIG. 38  is a schematic sectional perspective view of an embodiment of the present invention; 
         FIG. 39  is a schematic sectional view of the embodiment of  FIG. 38 ; 
         FIG. 40  is a sectional view along line DD in  FIG. 39 ; and 
         FIG. 41  is a graph showing a comparison of piston force versus force converted to work. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     In the following description, like reference characters designate like or corresponding parts throughout the several views. Also in the following description, it is to be understood that such terms as “forward”, “rearward”, “left”, “right”, “upwardly”, “downwardly”, and the like, are words of convenience and are not to be construed as limiting terms. 
     One objective is to increase gas mileage of combustion engines by changing how the piston force is harnessed to accomplish work. This can be done by doing the following during the power stroke:
         Use the majority of the available piston force to do work by efficiently transferring the piston force close to TDC where the pressure is at its maximum (see  FIG. 5 ).   Maximize the force peak by igniting the fuel mixture so that the pressure peak occurs close to TDC.   Use geometry to maintain the piston force peak longer while rotating the crank.   Since torque is equal to force times crank radius, use a larger crank to create more torque from the available piston force.       

       FIG. 6  shows a force diagram of an embodiment of the machine of the present invention. This embodiment uses a crank and connecting rod, as does the crank slider mechanism shown in  FIG. 1 , but introduces a multi-linkage connecting rod mechanism and a force transfer mechanism used to efficiently transfer force when the crank angle is less than 45 degrees, when the pressure is at its maximum and the conventional crank slider mechanism is least efficient. The force transfer mechanism can have different configurations, including a non-circular gear mechanism and a linkage mechanism.  FIG. 6  shows a force diagram with the linkage mechanism. For analysis in this document, an elliptical gear is used as the non-circular gear mechanism since its shape is easily defined. 
     The multi-linkage connecting rod mechanism (“MLCR”) includes additional hinge links between the connecting rod and the crankshaft. This change allows the main connecting rod to descend in relation to the crank. The need for this can be visualized by looking at  FIG. 6 . If the hinge links were not present and α=β=0, then the crank would not allow the connecting rod to transfer force to the force transfer mechanism, stopping the transfer of force used to turn the crank. 
     In addition, the piston is no longer on the same centerline as the crank, as compared to a conventional crank slider mechanism. Offsetting the centerline provides additional space for the hinge. While an offset centerline provides advantages, the present machine will also work with a piston and crank on the same centerline, if required for a specific application. Different approaches can be used to reset the geometry of the multi-linkage connecting rod mechanism and the force transfer mechanism, as will be discussed in further detail below. 
     The simple linkage force transfer mechanism of  FIG. 6  will first be analyzed. By using a static force analysis and summing the horizontal and vertical forces, the relationship between piston force and force harnessed to turn the crank can be determined for this embodiment: 
         FC=FP *Cos(β)
 
         FC *Cos(β)= F   W *Cos(90− a )+ FL *Cos(μ)
 
         FC *Sin(β)+ FL *Sin(μ)= F   W *Sin(90− a )
 
     Evaluating the relationship for the initial condition when a=β=0 and for μ equal to 45 degrees shows that F W  is equal to 50% of piston force FP for this initial condition. This is the initial transfer rate and will increase dramatically as the crank is rotated. The simple linkage force transfer mechanism is the least efficient at harnessing piston force of the embodiments described herein but has the advantage of being the least costly to implement. Variations of this embodiment are possible. One has a force transfer mechanism having a fixed pivot point. One has a force transfer mechanism having a moveable pivot point. 
     A force diagram for a non-circular gear force transfer mechanism is shown in  FIG. 7 . Comparing  FIG. 7  to  FIG. 6 , it can be seen that the simple linkage force transfer mechanism of  FIG. 6  has been replaced with a non-circular gear mechanism in  FIG. 7 . With the non-circular gear mechanism, the connecting rod force FC is transferred to the non-circular gear causing it to turn and as the non-circular gear turns its radius changes, creating force FE. The percentage of force transferred from FC to FE depends on multiple aspects, including the relationship between the major and minor radiuses of the non-circular gear, the gear tooth profile, the pivot point offset from the center of the non-circular gear, and the angle between the connecting rod and the non-circular major axis. When the piston is close to TDC, a non-circular gear, along with the multi-linkage connecting rod mechanism, is able to transfer a majority of the vertical piston force into horizontal force, thereby harnessing the piston&#39;s force at peak pressure. 
       FIG. 8  shows a first embodiment. The engine  10  has a casing  11  which includes a cylinder  12  and a crankshaft support  13 . The casing  11  can have a unitary construction or include a number of separate components attached together, such as a crankcase and cylinder. The cylinder  12  has a bore  14  in which a piston  16  is slidingly reciprocally disposed. A crankshaft  18  includes a main bearing journal  20  rotationally supported in the crankshaft support  13  and a crankpin  22  radially offset from an axis of the main bearing journal. A crank web  24  connects the main bearing journal  20  and the crankpin  22 . The engine includes other components as would be known to a person of ordinary skill in the art and those components are not described herein. Also, the embodiments disclosed below use many of the same components, or similar components as this embodiment. Therefore, repeat descriptions of components or portions of components may be omitted below when the components or portions of components are the same or similar to components or portions of components already described herein. 
     A multi-linkage connecting rod mechanism  26  (“MLCR”, see also  FIG. 9 ) includes a connecting rod  28  having a first end  30  connected to the piston  16  and a second end  32  opposite the first end  30 , the connecting rod  28  for reciprocating with the piston  16 . A first hinge link  34  includes a first end  36  pivotally connected to the connecting rod second end  32  and a second end  38  opposite the first hinge link first end  36 . A crankpin link  40  includes a first end  42  and a second end  44  rotationally connected to the crankpin  22 . In one embodiment, the multi-linkage connecting rod mechanism  26  also includes a second hinge link  46  with a first end  48  pivotally connected to the first hinge link second end  38  and a second end  50  pivotally connected to a first end  54  of a third hinge link  52 . The third hinge link  52  also includes a second end  56  pivotally connected to the crankpin link first end  42 . 
     Alternatively, the crankpin link first end  42  can be pivotally connected directly to the first hinge link second end  38 , and the second hinge link  46  and third hinge link  52  omitted. The crankpin link first end  42  can also be pivotally connected directly to the second hinge link  46  and the third hinge link  52  omitted. Any number of hinge links can be used between the connecting rod  28  and the crankpin link  40 . 
     The pivoting connections between the connecting rod, hinge links and crankpin link can have alternative constructions. In one embodiment, the connecting rod  28  has a unitary second end  32  which is inserted in a slot of the first hinge link first end  36 . The slot can be formed from a fork or clevis construction of the first hinge link and/or can be formed by separate portions (or halves) of the first hinge link sandwiching the connecting rod second end  32 . The components can also have a layered construction and be assembled in an interleaving manner with respect to adjacent components. Respective pins  58 ,  60 ,  62  and  64  engage between bores in the respective portions of the connecting rod, hinge links and crankpin link to provide the pivotal connections thereto. 
     The first hinge link  34  also includes a stop face  66  for engaging a stop face  68  of the third hinge link  52 . These stop faces  66  and  68  can engage one another when the first hinge link  34  pivots in one direction with respect to the third hinge link  52  to provide a positive stop to the extent of such pivoting and prevent further pivoting in that direction. The second hinge link  46  also includes a stop face  70  for engaging a stop face  72  of the crankpin link  40  to provide a positive stop to the pivoting of those components with respect to one another. There can also be counterpart engaging stop faces on the connecting rod and in other positions of the other components of the multi-linkage connecting rod mechanism to also limit pivoting in a direction and can have counterpart stop faces on an opposite side of the components to limit pivoting in the opposite direction as well. Although the stop faces provide a limit to the pivoting, in operation, the MLCR need not pivot to the maximum limit. The total range of pivoting of the connecting rod  28  with respect to the crankpin link  40  will usually be within a range of approximately 0-90° in either direction, as measured between longitudinal axes of the connecting rod and the crankpin link. An amount of pivoting allowed in one direction can be different from an amount of pivoting allowed in the opposite direction. The amount of pivoting between one component and another can be different than an amount of pivoting between one of those components and another component, or between two other components. 
     The engine also includes a force transfer mechanism  74  (see  FIG. 8 ) connecting the multi-linkage connecting rod mechanism  26  to the casing  11  for transferring a vertical piston force into a horizontal crankpin force. The force transfer mechanism  74  includes a pivot link  76  including a first end  78  pivotally connected to the casing  11  at fixed/stationary pivot point  79  and a second end  80  pivotally connected to a third pivot connection  82  of the first hinge link  34 . Alternatively, the connecting rod  28  or other components of the multi-linkage connecting rod mechanism can have the third pivot connection for connection to the pivot link second end  80 . The third pivot connection  82  is positioned on a major thrust side (on the left in  FIG. 8 ) of the first hinge link  34  and the pivot link first end  78  is pivotally connected to the casing  11  on the major thrust side of the first hinge link  34 . The pivot connections can be provided in any manner, using pins, bolts other threaded fasteners to provide the pivot mount, and can use bearings and/or bushings between the pivot mount and the pivoting component to reduce friction and wear. 
       FIGS. 10A-10F  show a sequence of six positions for the engine  10  as the engine rotates through a revolution.  FIG. 10A  shows the power stroke of the engine  10  close to TDC. The initial pivot angle between the connecting rod  28  and the crankpin link  40  is 45 degrees, which will harness 50% of the available piston force. As the piston  16  is pushed down during the power stroke, the piston force is transferred to the MLCR  26  while the force transfer mechanism  74  transfers the piston force to the crankpin  22  which turns the crankshaft  18 . Thus, the downward piston force is more rapidly converted to a sideways force acting on the crankpin  22  at this important initial stage of the power stroke.  FIGS. 10B-10F  show the changing shape of the MLCR  26  during the down stroke and then the raising of the piston and the resetting of the force transfer mechanism  74  and MLCR  26  during the up stroke. During the down stroke, the MLCR  26  will straighten out as the crank rotates. See  FIG. 10C  and compare with  FIGS. 10A and 10B .  FIG. 10D  shows that during the initial phase of the up stroke of the crankshaft  18 , the piston  16  continues to descend as the angle of the MLCR  26  changes. 
     As the crankshaft rotates past 180 degrees of rotation and the MLCR approaches horizontal at the end of the power stroke ( FIGS. 10C-10D ), the pivot link  76  can obstruct the path of the MLCR  26 . Adding the pivot connection at pin  58  between the connecting rod  28  and the first hinge link  34  allows the MLCR  26  to bend around the pivot link  76  ( FIG. 10D ). 
     As the crankshaft  18  continues rotation, the piston  16  will start to rise again ( FIG. 10E ) with the crankpin link  40  or other portion of the MLCR  26  contacting the pivot link  76  and causing it to rise to return to its default angle at TDC.  FIG. 10F  shows the resetting of the force transfer mechanism  74  and MLCR  26 . As the engine moves from the position in  FIG. 10F  to the position in  FIG. 10A , the force transfer mechanism  74  and MLCR  26  stops the upward motion of the piston  26  to prevent contact of the piston  16  with a head portion of the cylinder  12  and to reverse the direction of movement of the piston  16  for the down stroke. 
     While this example and the examples below are discussed in terms of engines, and particularly internal combustion reciprocating piston engines, the invention is applicable to other types of reciprocating piston fluid flow machines. 
       FIG. 11  shows a schematic view of a further engine  84  and  FIG. 12  shows a schematic detail view of a multi-linkage connecting rod mechanism  86  of the embodiment of  FIG. 11 . The multi-linkage connecting rod mechanism  86  (“MLCR”) includes a connecting rod  28  having a first end  30  connected to the piston  16  and a second end  32  opposite the first end  30 , the connecting rod  28  for reciprocating with the piston  16 . 
     A first hinge link  88  includes a first end  90  ( FIG. 12 ) pivotally connected to the connecting rod second end  32  and a second end  92  opposite the first hinge link first end  90 . The second end  92  is pivotally connected to the second hinge link first end  48  (previously shown in  FIG. 9 ). The third pivot connection  94  includes a first gear portion  96 . The force transfer mechanism  111  includes pivot link  112  having first end  78  and second end  114 . Pivot link second end  114  includes a second gear portion  116  engaging the first gear portion  96  in a meshing connection that allows the first hinge link  88  to pivot with respect to the pivot link  112  as the connecting rod  28  reciprocates. The first gear portion  96  is a rack gear and the second gear portion  116  is a curved non-circular gear having a different radius in an x-axis than a y-axis, although other gear arrangements can be used. Positioning of the gear portions, as well as other components of the various embodiments can also be reversed without departing from the scope of the invention. 
     A retention mechanism  98  is connected between the first hinge link  88  and the pivot link  112  to maintain the first gear portion  96  in meshing engagement with the second gear portion  116  throughout a range of pivoting of the multi-linkage connecting rod mechanism  86  with respect to the pivot link  112 . The retention mechanism  98  includes two side plates  100 , each having a first end  102  and a second end  104 . Each side plate  100  is fixedly attached on the first end  102  to the MLCR  86  (here the first hinge link  88 ) with side plate connectors  110 , which can be pins, threaded connections, cast or welded connections or other types of connections. Each side plate  100  includes a first slot  106  and a second slot  108  positioned on the second end  104 . Pivot link  112  includes a pair of retention bores  118  for securing retention pins  120  (see  FIG. 11 ). The retention pins  120  and  122  engage the first and second slots  106  and  108 , respectively, to maintain the curved non-circular gear in meshing engagement with the rack gear throughout a range of pivoting of the multi-linkage connecting rod mechanism with respect to the pivot link. The shape of the slots is configured to provide the desired meshing engagement between the gear portions as they move through their cycles. The retention pins  120  and  122  are located at different points on the pivot link  112  and follow different paths during gear rotation. This prevents separation of the non-circular gear  116  from the rack gear  96  as at least of the pins  120  and  122  limits the movement of the two gears away from one another. 
     In this embodiment, as can be seen in  FIG. 11 , the piston center line is brought closer in line with the crankshaft center line, and in one variation, is aligned with the crankshaft centerline. This is different from the fixed pivot point link force transfer alternative of  FIGS. 8-10 . By lining up the crankshaft and piston center lines, the engine  84  is able to be narrower and the MLCR  86  is not required to bend at extreme angles during the up stroke to get around the force transfer mechanism  111 . A disadvantage of aligning the center lines is that the MLCR  86  at TDC is bent slightly (See  FIG. 14B ) which reduces force transfer efficiency close to TDC compared to optimal efficiency. However, even with this disadvantage, this alternative is still more efficient that the embodiment of  FIGS. 8-10 . 
       FIG. 13A-13F  show a sequence of six positions for the engine  84  as the engine  84  rotates through a revolution, including the changing shape of the MLCR  86  during the down stroke, and the raising of the piston and the resetting of the force transfer mechanism  111  and MLCR  86  during the up stroke. During the down stroke, the MLCR  86  will straighten out as the crankshaft  18  rotates.  FIGS. 14C and 14D  show that this embodiment does not cause the piston  16  to drop during the initial stages of the up stroke, as the previous embodiment of  FIGS. 8-10  does. In addition, this alternative reduces piston stroke for a fixed crank stroke compared to the previous embodiment. As the engine moves between the positions of  FIGS. 14F and 14A , when the piston  16  approaches TDC, the force transfer mechanism  111  including non-circular gear  116 , retention mechanism  98 , and MLCR  86  stop the upward motion of the piston  16 .  FIG. 14A  shows the engine  84  close to TDC. During the power stroke, the piston force is transferred to the MLCR  86 , which in combination with the force transfer mechanism  111 , transfers the force to the crank pin which turns the crank. 
     The above discussed fixed pivot point alternatives are the simplest to implement and require less reciprocating mass than sliding pivot point alternatives. The drawback of these fixed pivot point alternatives is that they can require difficult connecting rod angles during the up stroke to move past the force transfer mechanism. These difficult angles require more work to raise the piston and can cause the piston stroke to be longer than that for sliding pivot point alternatives discussed below. 
     In alternative embodiments, the engine uses a sliding pivot point, as opposed to the alternatives discussed above using the fixed pivot point  79 . This allows the pivot point to move during the up stroke so that the force transfer mechanism is moved out of the way of the MLCR. This movement allows the MLCR to maintain shallow angles during the up stroke, reducing the required force to move the piston. The shallower angles can allow the MLCR to bend in only one direction. This has the advantage of a larger crank stroke than the piston stroke, as the changing shape of the MLCR can be taken advantage of during the down stroke. In addition, during 360 degrees of crankshaft rotation, the piston  16  spends more time close to TDC during the up stroke compared to the fixed pivot point alternatives or a conventional crank slider mechanism. The more time spent close to TDC during the compression stroke allows more time for the fuel mixture to burn. This allows the fuel mixture to reach a higher temperature earlier in the combustion process, which increases the pressure peak, as the volume is smallest in the early phases of the power stroke. 
     In embodiments disclosed below, each of the sliding pivot point alternatives includes three components. The first is a sliding mechanism used to control the path the pivot point takes when it is moving. The second is a locking mechanism used to lock the sliding mechanism into position during the down stroke; the last component is a motion control mechanism used to ensure that the links of the MLCR stay locked during the up stroke. There are many ways to lock the pivot point and control its motion during movement. Alternatives are described later below. 
       FIG. 15  shows a schematic sectional view of an engine incorporating a sliding pivot point in the force transfer mechanism. Engine  124  includes an MLCR  126  and a force transfer mechanism  128 . The MLCR  126  is similar to the MLCR  26  of  FIGS. 8-10 , but in this embodiment, the third pivot connection  130  is provided on the connecting rod  132 . Also, this embodiment uses only one hinge link  134  between the connecting rod  132  and the crankpin link  40 . See  FIG. 17 . 
     A moving pivot point mechanism  136  connects the pivot link first end  78  to the casing  11  (see  FIG. 17 ). The moving pivot point mechanism  136  includes a movable mount  138 , the pivot link first end  78  having a pivotal connection  140  with the movable mount  138 . The movable mount  138  is shown as having a generally rectangular block body  142  (see also  FIG. 16 ) for being received in and sliding in a correspondingly shaped guide path  144  of guide path device  145 , the guide path  144  providing a defined path along which the movable mount  138  can travel. The body  142  includes a pivot bore  146  for receiving a shaft for the pivot  140  and a locking detent  148  for engaging a locking pin  150 . The movable body also includes a bore  152  used for locking the pivot shaft to the moving pivot point. Although only one movable mount  138  can be used, in a preferred embodiment, two guide paths  144  are provided for supporting two movable mounts  138 , one on each side of the casing  11 , that is, on opposite sides of the pivot link  76 , and the pivot  140  runs between the two pivot bores  146  with the pivot link positioned therebetween. This creates a strong connection because the shaft is supported at both ends and locks the shaft in place between the opposite sides of the casing  11 . 
     A locking mechanism  154  connects with the movable mount  138  for locking the movable mount  138  in a locked position, as shown in  FIG. 17 . The locking mechanism  154  includes the locking pin  150  being driven by the crankshaft  156  and having locking portion  151  to engage the locking detent  148  of the movable mount  138  when the movable mount  138  is moved to the locked position to lock the movable mount  138  in the locked position. The locking pin  150  is also driven by the crankshaft to disengage from the locking detent  148  when the piston  16  is on a downstroke. Crankshaft  156  includes a first cam mechanism  158  for controlling motion of the movable mount  138 , as will be discussed further below and a second cam mechanism  162  for controlling the locking mechanism. Although only one of each of the cam mechanisms  158  and  162  need be provided on the crankshaft, in the embodiment shown, one each of the cam mechanisms  158  and  162  is provided on each side of the crankpin  22  so that each mechanism can operate an individual locking/motion control mechanism positioned on respective sides of the crankpin  22 . The cam mechanisms can be machined into the crankshaft or can be separate components attached to the crankshaft, either removably or permanently. A separate camshaft can also be provided, driven by the crankshaft. 
     The locking mechanism  154  further includes the second cam mechanism  162  driven by the crankshaft  156  and having a second cam path  164 . A second cam follower  166  engages between the second cam path  164  and the locking pin  150 . The second cam follower  166  (see also  FIG. 19 ) includes a cam engaging surface  167  and a ramped driving surface  168 . A spring  170  biases the second cam follower  166  in a direction toward the second cam path  164  and a spring can be used to bias the locking pin  150  in a direction away from the locking detent  148  when the locking pin  150  is not being driven into engagement with the locking detent  148  by the second cam follower  166 . Alternatively, the spring can be omitted and the shape of the pin and follower configured to move the cam follower. 
     The locking pin  150  includes a driven surface  172  for engaging the second cam follower driving surface  168  such that movement of the second cam follower  166  in a locking direction causes the locking pin  150  to move toward engagement with the locking detent  148  of the movable mount  138  and movement of the second cam follower  166  away from the locking direction allows the locking pin  150  to move away from engagement with the locking detent  148  of the movable mount  138 . When the locking pin  150  engages the locking detent  148 , the movable mount  138  is prevented from moving from the locked position. When the locking pin  150  is disengaged from the locking detent  148 , the movable mount  138  is allowed to move along the guide path  144 , subject to control by a motion control mechanism  172 . 
     The motion control mechanism  172  for controlling movement of the movable mount  138  includes the first cam mechanism  162  driven by the crankshaft  156  and having the first cam path  160 . A motion control linkage  180  operatively connects between the first cam path  160  and the movable mount  138  to allow the first cam path  160  to control movement of the movable mount  138 . The first cam path  160  includes a first portion constructed and arranged to allow the movable mount  138  to move along the guide path  144  in a direction away from the MLCR  126  on a first portion of an upstroke of the piston  16 , where the movement of the connecting rod  132  and pivot link  76  drive the movable mount  138  away from the MLCR  126  and crankshaft  156 . 
     The first cam path  160  also includes a second portion constructed and arranged to move the movable mount  138  along the guide path  144  in a direction toward the MLCR  126  and crankshaft  156  on a second portion of an upstroke of the piston  16  until reaching the locked position. On this second portion of the first cam path  160 , the first cam path  160  can drive the motion control linkage  180  to drive the movable mount  138  toward the locked position. The motion control linkage  180  further includes a first cam follower engaging the first cam path, the first cam follower  182  including a driving surface  184 . A rocker arm  186  is pivotally mounted to the casing  11  with pivot  188  and has a first end  190  connecting with the movable mount  138  and a second end  192  opposite the first end  190  connecting with the first cam follower driving surface  184 . The rocker arm  186  thus transfers motion from the first cam path  160  and the first cam follower  182  to the movable mount  138 . In an alternative embodiment, a biasing force to bias the movable mount  138  toward the locking position can be provided by a spring, hydraulically, by a solenoid or electric motor or by another mechanism. 
     A cam follower housing  194  (see  FIG. 18 ) is connected to the casing  11  and includes a first cam follower slot  196  slidably receiving the first cam follower  182 , a second cam follower slot  198  slidably receiving the second cam follower  166  and a locking pin slot  200  slidably receiving the locking pin  150 . 
     The guide path device guide path  144  includes a straight slot portion  202  slidably receiving the movable mount  138 . In this embodiment, the first and second cam follower slots  196  and  198  are aligned radially with respect to an axis of the crankshaft  156  and the locking pin slot  200  is aligned normal to the second cam follower slot  198 . 
     This embodiment is similar to the fixed pivot point embodiment of  FIGS. 8-10  but the pivot point  140  is on a sliding movable mount  138 . The movable mount  138  slides out of the way of the MLCR  126  when pivot link  76  is horizontal as the crankshaft  156  rotates past 180 degrees. This eliminates the need for the third hinge link  52  and allows the hinge links of the MLCR  126  to rotate a smaller amount past center, where they lock into position at the bottom of the stroke. 
       FIGS. 22A-22F  show a sequence of six positions for the engine  124  as the engine  124  rotates through a revolution, including the changing shape of the MLCR  126  during the down stroke, and the raising of the piston  16  and the resetting of the force transfer mechanism  128  and MLCR  126  during the up stroke.  FIG. 22A  shows the engine  124  with the piston  16  at TDC. TDC for the piston of this engine  124  starts before the crankpin  22  is at the top. By doing this, the crankshaft  156  can be rotated without the piston  16  dropping as much as with a conventional crank slider mechanism would for the same amount of crankshaft rotation. This change allows the piston peak power to be maintained for longer, which increases the overall power level generated by the same amount of fuel burned. This is true for all sliding pivot point alternatives (further embodiments are discussed below). 
       FIGS. 22B and 22C  show the down stroke and the changing shape of the MLCR  126 . All of the moving pivot point alternatives are able to take advantage of the changing shape of the MLCR and that the MLCR bends substantially only in one direction to support crank strokes that are larger than the piston stroke. That is, the MLCR  126  can bend in a direction away from the force transfer mechanism, and return to a starting position, but the MLCR  126  does not substantially bend from a longitudinal axis of the connecting rod in a direction toward the force transfer mechanism. It may go over center in a direction toward the force transfer mechanism to a small amount but not to the same degree as in the other direction. It can also be limited to not bending beyond center in a direction toward the force transfer mechanism. 
     By the position of  FIG. 22C , almost all of the piston force has been harnessed to perform work. At this point in the stroke, the goal is to prepare for the next power stroke by lifting the piston  16  as the crankshaft  156  rotates past 180 degrees of rotation ( FIGS. 22D and 22E ). By the position of  FIG. 22D , the moving pivot point mechanism  136  has been unlocked, allowing the pivot link  76  to slide out of the way of the MLCR  126  as the crankshaft  156  continues to rotate. This Figure shows the hinge links of the MLCR  126  as being locked, where the major thrust side stop faces have engaged one another, to prevent further bending in that direction and to allow the piston  16  to be raised. In this embodiment, it is preferred that bending toward the moving pivot point mechanism  136  be limited to essentially a straight or nearly straight MLCR with up to approximately 10° bending toward the moving pivot point mechanism allowed. In the position of  FIG. 22F , the piston  16  continues to be raised and the moving pivot point mechanism  136  is being moved back to its starting position at TDC were it will be locked into place. After the moving pivot point mechanism is locked in place, the crankshaft  156  continues to rotate and the hinge links of the MLCR will bend to the other side, away from the moving pivot point mechanism  136  back to their starting position ready for the next power stroke, Between the positions of  FIGS. 22F and 22A , the force transfer mechanism  128  and MLCR  126  interact to stop the upward motion of the piston  16  and to reverse its direction of movement for the down stroke. 
       FIG. 23  shows a variation of the embodiment of  FIGS. 15-22 , where the force transfer mechanism  209  is similar to the force transfer mechanism  74  of the embodiment of  FIGS. 11-14 . On this engine  207 , the first gear portion  210  is positioned on the connecting rod  212  of the MLCR  208  (see  FIG. 24 ), as opposed to being positioned on the first hinge link. As with the MLCR  126 , the moving pivot point mechanism  136  allows the MLCR  208  to have only connecting rod  212 , hinge link  134  and crankpin link  40 . 
       FIGS. 25A-25F  show a sequence of six positions for the engine  207  as the engine  207  rotates through a revolution, including the changing shape of the MLCR  208  during the down stroke, and the raising of the piston  16  and the resetting of the force transfer mechanism  209  and MLCR  208  during the up stroke. 
       FIG. 25A  shows the engine  207  with the piston  16  at TDC. TDC for the piston of this engine  207  starts before the crankpin  22  is at the top. By doing this, the crankshaft  156  can be rotated without the piston  16  dropping as much as with a conventional crank slider mechanism would for the same amount of crankshaft rotation. This change allows the piston peak power to be maintained for longer, which increases the overall power level generated by the same amount of fuel burned. This is true for all sliding pivot point alternatives (further embodiments are discussed below). 
       FIGS. 25B and 25C  show the down stroke and the changing shape of the MLCR  208 . All of the sliding pivot point alternatives are able to take advantage of the changing shape of the MLCR and that the MLCR bends substantially only in one direction to support crank strokes that are larger than the piston stroke. 
     By the position of  FIG. 25C , almost all of the piston force has been harnessed to perform work. At this point in the stroke, the goal is to prepare for the next power stroke by lifting the piston  16  as the crankshaft  156  rotates past 180 degrees of rotation (FIGS.  25 D and  25 E). By the position of  FIG. 25D , the moving pivot point mechanism  136  has been unlocked, allowing the pivot link  112  to slide out of the way of the MLCR  208  as the crankshaft  156  continues to rotate. This Figure shows the hinge links of the MLCR  208  as being locked, where the major thrust side stop faces have engaged one another, to prevent further bending in that direction and to allow the piston  16  to be raised. In this embodiment, it is preferred that bending toward the moving pivot point mechanism  136  be limited to essentially a straight or nearly straight MLCR. In the position of  FIG. 25F , the piston  16  continues to be raised and the moving pivot point mechanism  136  is being moved back to its starting position at TDC were it will be locked into place. After the moving pivot point mechanism is locked in place, the crankshaft  156  continues to rotate and the hinge links of the MLCR will bend to the other side, away from the moving pivot point mechanism  136  back to their starting position ready for the next power stroke. Between the positions of  FIGS. 25F and 25A , the force transfer mechanism  209  and MLCR  208  interact to stop the upward motion of the piston  16  and to reverse its direction of movement for the down stroke. 
       FIG. 26  shows a variation of the embodiment of  FIGS. 23-25 , using a different type of force transfer mechanism  215  including moving pivot point mechanism  216 , motion control mechanism  218  and locking mechanism  220  and MLCR  208 . On this engine  214 , the moving pivot point mechanism  216  does not travel along a linear path as in the previous embodiment but travels along a semi-circular path. Crankshaft  222  includes a first cam mechanism  224  having a first cam path  226  and a second cam mechanism  228  having a second cam path  230 . As with the embodiments above, while one each of the cam mechanisms and cam paths can be used, in a preferred embodiment, one each of the cam mechanisms and one each of the cam paths is used on each side of the crankpin  22  for strength. 
     The moving pivot point mechanism  216  includes a movable mount  232  having a pivotal connection  234  to the casing  11  to pivot about a movable mount axis  235  established by the pivotal connection  234 . A pivoting pivot link connection  236  is connected to the pivot link first end  78 , the pivot link connection  236  positioned radially outward of the movable mount axis  235  such that pivoting of the movable mount  232  about the movable mount axis  235  causes the pivot link connection  236  to move along an arcuate guide path  238 . The pivot link connection  236  can directly engage the arcuate guide path  238 . Alternatively, a guide element  237 , such as a roller bearing or bushing, connected to the pivot link connection  236  can engage the arcuate guide path  238 . Alternatively, there can be no engagement between the pivot link connection  236  and the arcuate guide path  238 . 
     The movable mount  232  is shown as having a partial disc configuration (See  FIG. 28 ) but can have a full disc or other configuration. It includes pivot bore  240  for the pivot link connection  236  and center bore  242  for the pivotal connection  234 , the center bore  242  establishing the movable mount axis  235 . A counterbalance portion  244  is positioned opposite the movable mount axis  235  from the pivot link connection  236 . The movable mount  232  includes a locking surface  246  for engaging a locking pin  248 . In this embodiment, the locking surface  246  is formed as a V-shaped notch or detent in an outer circumference of the counterbalance portion  244 , but it can have alternative configurations and positions. 
     A moving pivot point housing  250  (see  FIGS. 30 and 31 ) supports the movable mount  232  and other components of the moving pivot point mechanism  216 , motion control mechanism  218  and locking mechanism  220 . As seen in the first side view of  FIG. 31 , the moving pivot point housing  250  includes a recessed portion  252  for receiving the movable mount  232  and a boss portion  254  for providing the pivotal connection  234 . The movable mount  232  can pivot directly on boss portion  254  via center bore  242  or a bearing or bushing can be interposed between the boss portion  254  and center bore  242 . Alternative constructions can also be used including replacing boss portion  254  with a bore for receiving a pivot shaft to support the movable mount  232 . The moving pivot point housing  250  also includes a second cam follower slot  256  for slidably receiving a second cam follower  258  and a locking pin slot  260  for slidably receiving a locking pin  262 . See the second side view of the moving pivot point housing  250  in  FIG. 30 . A pocket portion  264  associated with the second cam follower slot is provided for a return spring to bias the second cam follower  258  toward the second cam mechanism  228 . Alternatively, the spring can be omitted and the shape of the pin and follower configured to move the cam follower. 
     The locking mechanism  220  includes the second cam mechanism  228  driven by the crankshaft  222  and having the second cam path  230 . The second cam follower  258  engages between the second cam path  230  and the locking pin  262 . The second cam follower  258  (see also  FIG. 33 ) includes a cam engaging surface  266  and a ramped driving surface  268 . A spring can bias the second cam follower  258  in a direction toward the second cam path  230  and a spring can bias the locking pin  248  in a direction away from the locking detent  246  when the locking pin  248  is not being driven into engagement with the locking detent  246  by the second cam follower  258 . Alternatively, the spring can be omitted and the shape of the pin and follower configured to move the cam follower. 
     The locking pin  262  (see  FIG. 32 ) includes a driven surface  274  for engaging the second cam follower driving surface  268  such that movement of the second cam follower  258  in a locking direction causes the locking pin  262  to move toward engagement with the locking detent  246  of the movable mount  232  and movement of the second cam follower  258  away from the locking direction allows the locking pin  262  to move away from engagement with the locking detent  246  of the movable mount  232 . Locking pin  262  also has engaging surface  276  for engaging the locking detent  246 . When the locking pin  262  engages the locking detent  246 , the movable mount  232  is prevented from moving from the locked position. When the locking pin  262  is disengaged from the locking detent  246 , the pivot link connection  236  is allowed to move along the arcuate guide path  238 , subject to control by the motion control mechanism  218 . 
     The motion control mechanism  218  for controlling movement of the movable mount  232  includes the first cam mechanism  224  driven by the crankshaft  222  and having the first cam path  226 . See  FIG. 34 . A motion control linkage  278  operatively connects between the first cam path  226  and the movable mount  232  to allow the first cam path  226  to control movement of the movable mount  232 . The first cam path  226  includes a first portion constructed and arranged to allow the pivot link connection  236  of movable mount  232  to move along the arcuate guide path  238  in a direction away from the MLCR  208  on a first portion of an upstroke of the piston  16 , where the movement of the connecting rod  212  and pivot link  112  drive the pivot link connection  236  away from the MLCR  208  and crankshaft  222 . 
     The first cam path  226  also includes a second portion constructed and arranged to move the pivot link connection  236  along the arcuate guide path  238  in a direction toward the MLCR  208  and crankshaft  222  on a second portion of an upstroke of the piston  16  until reaching the locked position. On this second portion of the first cam path  226 , the first cam path  226  can drive the motion control linkage  278  to drive the pivot link connection  236  toward the locked position. The motion control linkage  278  further includes a first cam follower  280  engaging the first cam path  226 , the first cam follower  280  including a driving surface  282 . The first cam follower  280  is slidably positioned in a first cam follower slot  281  in one or both of the moving pivot point housing  250  or cam follower housing  292 . A rocker arm  284  is pivotally mounted to the casing  11  with pivot mount  286  and has a first end  290  connecting with the pivot link connection  236  and a second end  288  opposite the first end  290  connecting with the first cam follower driving surface  282 . The rocker arm  284  thus transfers motion from the first cam path  226  and the first cam follower  280  to the pivot link connection  236 . In an alternative embodiment, a biasing force to bias the pivot link connection  236  toward the locking position can be provided by a spring, hydraulically, by a solenoid or electric motor or by another mechanism. A cam follower housing  292  is connected to the moving pivot point housing  250  to cover the first cam follower  280 . In this embodiment, the first and second cam follower slots  281  and  256  are aligned radially with respect to an axis of the crankshaft  222  and the locking pin slot  260  is aligned normal to the second cam follower slot  256 . The rocker arm first end  290  can engage the pivot link connection  236 , the guide element  237 , the movable mount  232 , the pivot link  112  or other structure to control movement of the movable mount  232 . The cam can also operate directly on the rocker arm, as the follower, or on the pivot mount. 
     The moving pivot point housing  250  can be a separate component attached to the casing  11 , can be integrally provided with the casing  11  or a combination of both. 
     In one embodiment, generally duplicate, or mirror image, moving pivot point mechanisms are provided on each side of the crankpin with a first single shaft running therebetween to support the pivotal connection  234  and a second single shaft also running therebetween to support the pivot link connection. This provides a strong structure because the respective shafts are each supported at two separated ends. 
     This embodiment has the advantage of being easier to balance and requiring less space as compared to the linear movable mounts discussed above. 
       FIGS. 35A-35F  show a sequence of six positions for the engine  214  as the engine  214  rotates through a revolution, including the changing shape of the MLCR  208  during the down stroke, and the raising of the piston  16  and the resetting of the force transfer mechanism  215  and MLCR  208  during the up stroke. 
       FIG. 35A  shows the engine  214  with the piston  16  at TDC. TDC for the piston of this engine  214  starts before the crankpin  22  is at the top. By doing this, the crankshaft  222  can be rotated without the piston  16  dropping as much as with a conventional crank slider mechanism would for the same amount of crankshaft rotation. This change allows the piston peak power to be maintained for longer, which increases the overall power level generated by the same amount of fuel burned. 
       FIGS. 35B and 35C  show the down stroke and the changing shape of the MLCR  208 . All of the sliding pivot point alternatives are able to take advantage of the changing shape of the MLCR and that the MLCR bends substantially only in one direction to support crank strokes that are larger than the piston stroke. 
     By the position of  FIG. 35C , almost all of the piston force has been harnessed to perform work. At this point in the stroke, the goal is to prepare for the next power stroke by lifting the piston  16  as the crankshaft  222  rotates past 180 degrees of rotation ( FIGS. 35D and 35E ). By the position of  FIG. 35D , the moving pivot point mechanism  216  has been unlocked, allowing the pivot link  112  to slide out of the way of the MLCR  208  as the crankshaft  222  continues to rotate.  FIG. 35D  shows the hinge links of the MLCR  208  as being locked, where the major thrust side stop faces have engaged one another, to prevent further bending in that direction and to allow the piston  16  to be raised. In this embodiment, it is preferred that bending toward the moving pivot point mechanism  216  be limited to essentially a straight or nearly straight MLCR. In the position of  FIG. 35F , the piston  16  continues to be raised and the moving pivot point mechanism  216  is being moved back to its starting position at TDC were it will be locked into place. After the moving pivot point mechanism  216  is locked in place, the crankshaft  222  continues to rotate and the hinge links of the MLCR will bend to the other side, away from the moving pivot point mechanism  216  back to their starting position ready for the next power stroke. Between the positions of  FIGS. 35F and 35A , the force transfer mechanism  215  and MLCR  208  interact to stop the upward motion of the piston  16  and to reverse its direction of movement for the down stroke. 
     While the embodiments described herein having moving pivot point mechanisms are shown having locking mechanisms, it is also contemplated that the locking mechanisms can be omitted with locking control provided by the motion control mechanism. That is, the motion control mechanism prevents substantive movement of the moving pivot point in the locked position. In such embodiments where the locking mechanism is omitted, the motion control mechanism can be made stronger to provide the necessary locking force. 
     The embodiments shown herein with a moving pivot are shown using linear or arcuate/circular movement. However, the moving pivot need not be so limited and can use any type of movement and mechanism to move out of the way of the MLCR. 
       FIGS. 36 and 37  show a variation of the embodiment of  FIGS. 23-25 . In this variation, engine  296  uses a different retention mechanism for maintaining the first gear portion  298  of connecting rod  300  of MLCR  301  in meshing engagement with the second gear portion  116  of pivot link  112 . Retention mechanism  302  includes an idler roller  304  rotatably mounted on a slider  306 . Slider  306  is slidably positioned in a slider slot  308  connected to the casing  11 . The slider slot  308  may be integrally provided as part of the casing  11  or can be a separate component attached to the casing  11 . The slider can slide along a linear path as shown or can move along a different path. The slider can also have a different configuration and need not actually side in its motion. Rather, it could rotate or have another type of movement. The idler roller  304  is biased against a side of the connecting rod  300  to maintain the first gear portion  298  in meshing engagement with the second gear portion  116 . As the connecting rod  300  moves up and down and back and forth, the sliding action of the slider maintains the idler roller in engagement with the connecting rod  300 . The idler roller  304  can be biased directly or via biasing the slider  306 . The biasing can be provided by a spring, hydraulically, by a solenoid or electric motor or by another mechanism. The biasing element can be provided in the slider slot  308  or external to the slider slot  308 . Alternatively, the idler roller  304  can be mounted on a pivot arm attached to the casing  11  and the pivoting of the arm can maintain the idler roller in contact with the connecting rod  300 . The pivot link  112  can also be biased into engagement with the connecting rod  300  by a spring, hydraulically or otherwise to assist in maintaining the first gear portion  298  in meshing engagement with the second gear portion  116 . 
     This variation can also further include a piston control mechanism  310  which prevents the piston  16  from traveling above a designed piston TDC. See  FIGS. 38-40 . The piston control mechanism  310  includes a control button  312  attached to the connecting rod  309  for engaging a control recess  314  connected to the casing  11 . The engagement between the control button  312  and the control recess  314  defines a path of travel of the control button  312 , and thus the connecting rod  309 , at least on an upstroke of the piston  16  and connecting rod  309 . That is, the control recess  314  defines an outer path that the control button  312  can travel when engaged in the control recess  314  and the control recess  314  can be configured to limit the upward travel of the piston  16  as it approaches TDC to prevent over-excursion of the piston  16 . The piston control mechanism  310  can also include a control island  316  positioned in control recess  314 . The control island  316  can interact with the control button  312  to change an angle of the hinge links of the MLCR as it is raised, bending the hinge links into position for the next power stroke. The control button  312  can be fixed with respect to the connecting rod  309 , or can also be in the form of a roller such that it can roll as it travels along the control recess  314 . 
     As with other components discussed herein, the piston control mechanism  310  can be provided as a single unit, but it is preferable in certain situations to provide one each of the mechanisms  310  on opposite sides of the crankpin  22  to provide strength and support. 
     The various embodiments discussed above provide engine performance gains when compared to conventional crank slider mechanisms. There are generally three aspects of the performance gains: torque generated, efficiency of harnessing the piston force to do work, and time spent close to TDC during the compression stroke. 
     The torque generated during the power stroke for the Invention is larger than that of a crank slider mechanism given the same piston force and efficiency of harnessing this force to do work. This is because an engine as disclosed herein will have a larger crank than a comparative crank slider mechanism for the same drop in piston length from TDC to BDC. How much larger depends on many factors, including the geometry of the connecting rod and connecting rod gear. However, a typical figure would be ˜3 to 10% larger, which translates directly into more torque for the present engine as compared to a crank slider mechanism. 
     The efficiency of power transfer of the present engine depends on which alternative is used. The sliding pivot point non-circular gear alternatives are the most efficient and are used for the performance analysis in this section. For the non-circular gear alternatives, the efficiency depends on many parameters including the number of links in the variable geometry connecting rod, the major and minor radius of the connecting rod gear, offset of the pivot hole in the connecting rod gear, tooth profile of the connecting rod gear, initial angle of the connecting rod gear, placement of the connecting rod gear, and offset of the connecting rod compared to the crank center.  FIG. 41  shows a measured efficiency of a prototype sliding pivot point non-circular gear alternative as compared against a theoretical crank slider mechanism with a 4 inch stroke and a connecting rod 6 inches in length. This figure is the same as  FIG. 5  above, with the addition of the efficiency curve of the present engine. In  FIG. 41 , the “Pressure Available” curve is derived from the ideal gas law with temperature constant and  100  being the force available at TDC. The pressure available drops as the crank rotates and the volume of the combustion chamber expands. The “CSM curve” is derived from multiplying the crank slider mechanism percentage of force converted to work times the force available. The “Invention curve” is derived from multiplying the present engine percentage of force converted to work times the force available. This figure shows that the present engine is much more efficient that than the crank slider mechanism from TDC until the crank has rotated 60 degrees past TDC, where the piston force is 25% of its peak pressure. From 60 degrees onward the invention and crank slider mechanism have essentially equal efficiency. 
     For the sliding pivot point alternatives, during 360 degrees of crank rotation, the piston spends more time close to TDC during the compression stroke as compared to the fixed pivot point alternatives or a crank slider mechanism. The more time spent close to TDC during the compression stroke allows more time for the fuel mixture to burn. This allows the fuel mixture to reach a higher temperature earlier in the combustion process which increases the pressure peak as the volume is smallest in the early phases of the power stroke. 
     When these three advantages are looked at in totality, a significant gain of approximately 50 to 80% is seen in torque output compared to a crank slider mechanism. This is because the higher pressure level of the Invention is harnessed by its higher efficiency and acts upon a larger crank. Given the significant gain in torque output, a present engine will only need a fraction of the fuel burned to generate the same amount of torque as a crank slider mechanism engine. 
     For most engine applications, these advantages outweigh the higher complexity, cost, and reciprocating weight as compared to a crank slider mechanism and will result in very substantial fuel cost savings for each year of operation of the engine. 
     Various features of the various embodiments disclosed herein can be combined in different combinations to create new embodiments within the scope of the present invention. That is, one or more features from one or more of the embodiments can be combined with one or more features of one or more other embodiments to create new embodiments within the scope of the present invention. Any ranges given herein include any and all specific values within the range and any and all ranges within the given range. 
     The foregoing is a description of embodiments of the invention which are given here by way of example only. The invention is not to be taken as limited to any of the specific features as described, but comprehends all such variations thereof as come within the scope of the appended claims. 
     REFERENCE NUMERALS 
     
         
           10  engine 
           11  casing 
           12  cylinder 
           13  crankshaft support 
           14  bore 
           16  piston 
           18  crankshaft 
           20  main bearing journal 
           22  crankpin 
           24  crank web 
           26  multi-linkage connecting rod mechanism 
           28  connecting rod 
           30  connecting rod first end 
           32  connecting rod second end 
           34  first hinge link 
           36  first hinge link first end 
           38  first hinge link second end 
           40  crankpin link 
           42  crankpin link first end 
           44  crankpin link second end 
           46  second hinge link 
           48  second hinge link first end 
           50  second hinge link second end 
           52  third hinge link 
           54  third hinge link first end 
           56  third hinge link second end 
           58  pin 
           60  pin 
           62  pin 
           64  pin 
           66  first hinge link stop face 
           68  third hinge link stop face 
           70  second hinge link stop face 
           72  crankpin link stop face 
           74  force transfer mechanism 
           76  pivot link 
           78  pivot link first end 
           79  pivot point 
           80  pivot link second end 
           82  third pivot connection 
           84  engine 
           86  multi-linkage connecting rod mechanism 
           88  first hinge link 
           90  first hinge link first end 
           92  first hinge link second end 
           94  third pivot connection 
           96  first gear portion 
           98  retention mechanism 
           100  side plate 
           102  side plate first end 
           104  side plate second end 
           106  side plate first slot 
           108  side plate second slot 
           110  side plate connector 
           111  force transfer mechanism 
           112  pivot link 
           114  pivot link second end 
           116  second gear portion 
           118  retention bore 
           120  retention pin 
           122  retention pin 
           124  engine 
           126  multi-linkage connecting rod mechanism 
           128  force transfer mechanism 
           130  third pivot connection 
           132  connecting rod 
           134  hinge link 
           136  moving pivot point mechanism 
           138  movable mount 
           140  pivot 
           142  body 
           144  guide path 
           145  guide path device 
           146  pivot bore 
           148  locking detent 
           150  locking pin 
           151  locking pin locking portion 
           152  locking bore 
           153  driven surface 
           154  locking mechanism 
           156  crankshaft 
           158  first cam mechanism 
           160  first cam path 
           162  second cam mechanism 
           164  second cam path 
           166  second cam follower 
           167  second cam follower cam engaging surface 
           168  second cam follower driving surface 
           170  spring 
           172  motion control mechanism 
           180  motion control linkage 
           182  first cam follower 
           184  first cam follower driving surface 
           186  rocker arm 
           188  pivot 
           190  rocker arm first end 
           192  rocker arm second end 
           194  cam follower housing 
           196  first cam follower slot 
           198  second cam follower slot 
           200  locking pin slot 
           202  straight slot portion 
           207  engine 
           208  multi-linkage connecting rod mechanism 
           209  force transfer mechanism 
           210  first gear portion 
           212  connecting rod 
           214  engine 
           215  force transfer mechanism 
           216  moving pivot point mechanism 
           218  motion control mechanism 
           220  locking mechanism 
           222  crankshaft 
           224  first cam mechanism 
           226  first cam path 
           228  second cam mechanism 
           230  second cam path 
           232  movable mount 
           234  pivotal connection 
           235  movable mount axis 
           236  pivoting pivot link connection 
           237  guide element 
           238  arcuate path 
           240  pivot bore 
           242  center bore 
           244  counterbalance portion 
           246  locking surface 
           248  locking pin 
           250  moving pivot point housing 
           252  recessed portion 
           254  boss portion 
           256  second cam follower slot 
           258  second cam follower 
           260  locking pin slot 
           262  locking pin 
           264  pocket portion 
           266  cam engaging surface 
           268  ramped driving surface 
           270  spring 
           274  locking pin driven surface 
           276  locking pin engaging surface 
           278  motion control linkage 
           280  first cam follower 
           281  first cam follower slot 
           282  first cam follower driving surface 
           284  rocker arm 
           286  pivot mount 
           288  rocker arm first end 
           290  rocker arm second end 
           292  cam follower housing 
           296  engine 
           298  first gear portion 
           300  connecting rod 
           301  multi-linkage connecting rod mechanism 
           302  retention mechanism 
           304  idler roller 
           306  slider 
           308  slider slot 
           309  connecting rod 
           310  piston control mechanism 
           312  control button 
           314  control recess 
           316  control island