Abstract:
A force transfer mechanism is provided for an internal-combustion engine. In the most preferred configuration, the force transfer mechanism comprises two inter-connected first class levers that are driven by four pistons. Each of the pistons is driven through its non-powered strokes by the action of the piston in its powered stroke on the first class levers. The force transfer mechanism also drives the crankshaft through a single connecting rod. The force transfer mechanism provides for less frictional loss, and greater efficiency, than a typical internal combustion engine.

Description:
RELATED APPLICATIONS 
     This application is a continuation of and claims the benefit of my copending U.S. patent application Ser. No. 11/627,617, filed Jan. 26, 2007 entitled FORCE TRANSFER MECHANISM FOR AN ENGINE, which issued as U.S. Pat. No. 7,481,188, which is in turn a divisional of U.S. patent application Ser. No. 11/303,479, also entitled FORCE TRANSFER MECHANISM FOR AN ENGINE, filed Dec. 16, 2005, and issued as U.S. Pat. No. 7,219,467, both of which are herein incorporated by reference. 
    
    
     FIELD OF THE INVENTION 
     This invention relates to engines, and more particularly, but not limited to four-stroke internal combustion engines. 
     BACKGROUND OF THE INVENTION 
     To appreciate the advantages of the present invention, it is important to understand various aspects of how a typical internal combustion (IC) engine works. 
     In a typical IC engine, where the cylinders are fixed with respect to the engine frame, the motion of the connecting rods create side forces on their corresponding pistons that push against the cylinder walls. A standard four-cylinder internal combustion engine comprises four pistons, a crankshaft, and four connecting rods, each having a “big end” and a “small end.” Each piston is connected to the crankshaft through a corresponding connecting rod. The “big end” of the connecting rod is connected to one of several “rod journals” on the crankshaft—also known as a “crank throw”—that is offset from the “main journals” of the crankshaft. The “small end” of the connecting rod is pivotally attached to the piston via a “wrist pin.” As the piston reciprocates, the angle of the connecting rod with respect to the cylinder&#39;s longitudinal dimension changes. While the angular orientation of the connecting rod with respect to the piston is other than zero degrees, the connecting rod creates a side force on the piston against the cylinder wall. The magnitude of the force varies in relation to the angular orientation, gas pressures, and inertia forces. 
     To distribute these sideways forces, stabilize the path of the piston, and address friction issues, pistons are typically made with piston skirts that travel with the pistons inside the cylinders. While the pistons and skirts are likely lubricated to perform their role effectively, the larger the piston skirt, the greater a cross-sectional area of oil is sheared by the piston as it reciprocates. While a piston skirt performs an important function, its use creates an energy loss and thus decreases mechanical efficiency. 
     Furthermore, in a typical IC engine, while there is energy delivered by each piston to an output load through the crankshaft, a significant amount of energy is transferred through the crankshaft from each piston performing a power stroke to the pistons that are going through any of the three-non-powered strokes. Each piston cycles through a sequence of four strokes—the intake stroke, the compression stroke, the power stroke, and the exhaust stroke. At any given time while the engine is running, there is one cylinder performing a power stroke, another cylinder performing an exhaust stroke, another performing an intake stroke, and another a compression stroke. Because work must be performed to move each of the three pistons on non-powered strokes, the energy necessary to move them must be delivered by the piston performing the power stroke (excluding energy stored through inertia). In a standard in-line four-cylinder engine, this energy is delivered through the crankshaft. 
     The use of the crankshaft in a typical IC engine to transfer these between-cylinder forces and connect the cylinders to a common load increases the strength, size and rigidity requirements of the crankshaft as well as the size and number of bearing journals. Because the crankshaft is performing the dual roles of (1) transferring energy between the cylinders and (2) connecting the cylinders to a common load, the standard engine configuration results in a loss of energy and decreases mechanical efficiency. 
     Furthermore, a typical four-cylinder IC engine has five main journals and four rod journals to accommodate the four connecting rods driven by the pistons. The cross-sectional oil shear area across each of these nine-or-more relatively large bearings traveling through 360 degrees of rotation, multiplied by the distance the bearings travel per crankshaft revolution, is significant, and results in a loss of energy and decreases mechanical efficiency. 
     SUMMARY OF THE INVENTION 
     Various embodiments of the present invention improve upon the standard in-line 4 cylinder engine in a variety of ways. It is believed that the most preferred embodiment results in the greatest number of improvements. 
     In the preferred embodiment, a multiple watt-linkage force transfer mechanism is provided. (In 1784, James Watt reportedly invented what is now commonly referred to as a “watt linkage,” which is a mechanism for converting circular motion into near-straight-line motion, and incorporated it into a steam engine.) The force transfer mechanism comprises two “bell cranks” that are used to drive a single crank through a watt linkage mechanism. Each bell crank, in turn, is driven (and drives; depending on the stroke) two pistons through a watt linkage mechanism. The watt linkages connected to the pistons enable the connection end of the piston to travel along substantially straight paths, significantly reducing side loads against the piston walls. This potentially eliminates the need for piston skirts, or at least reduces their necessary lengths. It also potentially eliminates the need for a wrist pin between the pistons and their connection to the remainder of the mechanism. 
     Also, all four pistons preferably drive a single connecting rod. This simplifies the crankshaft design and the corresponding strength and rigidity requirements for the crankshaft by reducing the necessary number of rod journals and main journals on the crankshaft. It also provides a more efficient means of transferring the between cylinder forces. In one embodiment, the crankshaft can simply comprise a single rod journal and two main journals. 
     Moreover, the between-cylinder forces (between the power-stroke piston and the three pistons in a non-power-stroke) are transferred through the force transfer mechanism rather than through the crankshaft journals. The connecting rod transmits only the force remaining after the other three pistons consume the force that they need. In other words, all of the force transferred by the connecting rod to the crankshaft is used to drive the crankshaft and mechanisms connected to it. 
     Because the various components of the force transfer mechanism rotate about straight-pin type pivots that are either fixed in place, or that move along relatively straight paths (as opposed to the circular path of a rod journal caused by a revolution of the crankshaft), it is believed that the diameter of the pivot pins of the force transfer mechanism can be made fairly small, relative to the standard diameter of a rod journal, in order to safely and reliably deliver a desired amount of power. 
     Furthermore, the pistons are most preferably arranged in what is referred to herein as a “collinear H-cross configuration”—defined in the detailed description—in order to balance the inertial masses within the mechanism, except part of the mass of the connecting rod. In this configuration, the motion of each piston and its links are balanced by an equal and opposite motion of an opposing collinear piston and its links on the opposite side of the force transfer mechanism. Furthermore, the inertial mass of each piston is balanced on the opposite side of the bell crank fulcrum by an equivalent piston. 
     In summary, the most preferred embodiment potentially offers the following advantages over a standard in-line four-cylinder IC engine: 
     (1) The nearly complete elimination of side loads on the pistons potentially eliminates the need for piston skirts—limiting the piston friction losses to those resulting from the seals. 
     (2) The total cross-sectional area of the oil being sheared, times the shear distance, per crankshaft revolution will potentially be a small fraction of the area times distance being sheared in the typical engine. 
     (3) Because all of the between-cylinder forces generated are transferred through pins which rotate only a small amount, rather than through journals that have to rotate 360 degrees, each piston&#39;s work is not diminished by a chain of piston skirts, crankpin journals, and main bearings. In this manner, the mechanism also reduces part of the “pumping losses” a typical IC engine suffers when running at less-than-full throttle, where the pressure of the atmosphere outside the cylinder works against the engine. 
     (4) The force transfer mechanism internally balances the inertial masses associated with the four pistons. 
     It is believed that one of the inventive aspects driving the most preferred embodiments of the invention (but not necessarily all of the embodiments of the invention) is a focus on reducing the product of the total cross sectional area of oil being sheared times the distance it travels, instead of simply thinking about decreasing friction through reductions in bearing size, materials, and reduced friction surfaces. 
     While the most preferred embodiment is believed to benefit from all of the aforementioned advantages, the invention is broad enough to encompass embodiments that do not appropriate all, some, or any of these cited advantages. The invention disclosed herein encompasses numerous different kinds of embodiments—including embodiments that have only a single watt linkage mechanism; embodiments that utilize no watt linkage mechanisms; embodiments that utilize several sets of force transfer mechanisms; force transfer mechanisms that utilize what is later described as a “non-collinear H-cross configuration”; and force transfer mechanisms in which the pistons are not arranged in parallel with each other. The scope of any given claim will be set forth by the claim language itself—although this specification will explicitly define certain claim terms. 
     These and many other embodiments and advantages of the invention will be readily apparent to those skilled in the art from the following detailed description taken in conjunction with the annexed sheets of drawings, which illustrate the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a side view of one embodiment of a force transfer mechanism according to the present invention. 
         FIG. 2  is a second view of the force transfer mechanism of  FIG. 1  with all of the pistons halfway through their respective strokes. 
         FIG. 3  is a third view of the force transfer mechanism of  FIG. 1  with the crankshaft turned 180 degrees from the position depicted in  FIG. 1 . 
         FIG. 4  is an orthogonal view of the mechanism of  FIG. 1 . 
         FIGS. 5A and 5B  are cross-sectional views of various connecting links in  FIG. 1 . 
         FIG. 6A  is an enlarged view of the outer arm of  FIG. 1 . 
         FIG. 6B  is an orthogonal view of the outer arm of  FIG. 6 . 
         FIG. 6C  is a perspective view of the outer arm of  FIG. 6 . 
         FIGS. 7-9  are schematic representations of one embodiment of a force transfer mechanism at different stages of piston travel. 
         FIG. 10  is a schematic representation of an alternative embodiment of the force transfer mechanism of the present invention. 
         FIG. 11  is a dimensional diagram of a watt linkage mechanism usable in the present invention. 
         FIG. 12  is one embodiment of an eight cylinder configuration with two force transfer mechanisms disposed on opposite sides of the crank shaft. 
         FIG. 13  is one embodiment of a twelve cylinder configuration with three force transfer mechanisms stacked next to each other on the same side of the crank shaft. 
         FIG. 14  is an alternative embodiment of a force transfer mechanism according to the present invention which employs a standard crank and slider mechanism to drive the bell cranks. 
         FIG. 15  is another alternative embodiment which employs curved piston cylinders to drive the bell cranks. 
         FIG. 16  is an alternative embodiment of a force transfer mechanism that utilizes watt linkage mechanisms to drive the bell cranks, but not to drive the crank shaft. 
         FIG. 17  is an exploded view of one embodiment of a linkage between two connecting rods of  FIGS. 16 and 18 . 
         FIG. 18  is an alternative embodiment of a force transfer mechanism that utilizes no watt linkages, but which still achieves some of the advantages of the present invention. 
     
    
    
     DETAILED DESCRIPTION 
     Before the subject invention is described further, it is to be understood that the invention is not limited to the particular embodiments of the invention described below or depicted in the drawings. Many modifications may be made to adapt or modify a depicted embodiment without departing from the objective, spirit and scope of the present invention Therefore, it should be understood that, unless otherwise specified, this invention is not to be limited to the specific details shown and described herein, and all such modifications are intended to be within the scope of the claims made herein. 
       FIGS. 1-3  are views of one embodiment of a force transfer mechanism  100  according to the present invention. Force transfer mechanism  100  comprises a three-arm bell crank  140  driven by pistons  190  and  192  and another three-arm bell crank  150  driven by pistons  194  and  196 .  FIG. 1  shows pistons  190  and  194  at top dead center and pistons  192  and  196  at bottom dead center.  FIG. 2  shows each of the pistons  190 ,  192 ,  194  and  196  at the midpoint of their travel.  FIG. 3  depicts pistons  190  and  194  at bottom dead center and pistons  192  and  196  at top dead center. 
     The three-arm bell crank  140  pivots about a fulcrum  142  that is fixed in the frame. Likewise, bell crank  150  pivots about a fulcrum  152  that is also fixed in the frame. Bell crank  140  includes three bell crank pins  132 ,  144 , and  146 . Likewise, bell crank  150  includes three bell crank pins  134 ,  154 , and  156 . A connecting link  130  joins bell crank  140  to bell crank  150 . A first end of the connecting link  130  pivots about the bell crank pin  132  and the second end of the connecting link  130  pivots about bell crank pin  134 . Unlike the fulcrums  142  and  152 , the bell crank pins  132  and  134  are not fixed with respect to the engine block. 
     The distance between the fulcrum  142  and the bell crank pin  132  is the same as the distance between the fulcrum  152  and the bell crank pin  134 . Therefore, the connecting link  130  and bell cranks  140  and  150  together comprise what is known as a “watt linkage.” The bell cranks  140  and  150  comprise the “side links” of this watt mechanism, and the connecting link  130  comprises the “coupler link” of this watt mechanism (see explanation of “side links” and “coupler links” in the description below). 
     A pivot joint  136  is mounted at the midpoint of the connecting link  130 —that is, it is mounted at the point that is halfway between bell crank pins  132  and  134 . What is commonly referred to as the “small end”  137  ( FIG. 5 ) of the connecting rod  120 —that is, the end of the connecting rod that is opposite the crankshaft end  122  of the connecting rod  120 —is pivotally attached to the pivot joint  136  of the connecting link  130 . What is commonly referred to as the “big end” of the connecting rod  120 —that is, the crankshaft end  122  of the connecting rod  120 —pivots about a rod journal  112  of the engine crankshaft  110 . Because the pivot joint  136  is at the midpoint of the connecting link  130 , which is the “coupler link” of a watt mechanism, the pivot joint  136  drives the connecting rod  120  along a substantially straight path. 
       FIG. 1  depicts not just one watt linkage but in fact five watt linkages. Watt linkages are used to transfer the force from each piston  190 ,  192 ,  194 , and  196  to the bell cranks  140  and  150 . Each piston  190 ,  192 ,  194 , and  196  is connected to a corresponding connecting link  160 ,  165 ,  170 , or  175  via a piston pin  162 ,  167 ,  172 , or  177  mounted at the midpoint of the connecting link  160 ,  165 ,  170 , or  175 . These connecting links  160 ,  165 ,  170 , and  175  each comprise a “coupler link” of a watt linkage mechanism. 
     Piston  190  drives connecting link  160 , which is pivotally mounted both to the bell crank  140  via bell crank pin  144 , and to the outer arm  180  via inner joint  164 . Outer arm  180 —which forms one of the “side links” of a watt linkage mechanism—is pivotally mounted in fixed relation to the engine block via an outermost bearing  181 . The bell crank  140  forms the other “side link” of this watt linkage mechanism. Because of this watt linkage mechanism, the connection end of piston  190  is able to travel in a substantially straight line as it drives the connecting link  160 . 
     Piston  192  drives connecting link  165 , which is pivotally mounted both to bell crank  140  via bell crank pin  146 , and to the outer arm  182  via inner joint  169 . Outer arm  182 —which forms one of the “side links” of another watt linkage mechanism—is also pivotally mounted in fixed relation to the engine block via outermost bearing  183 . The bell crank  140  forms the other “side link” of this watt linkage mechanism. Because of this watt linkage mechanism, the connection end of piston  192  is able to travel in a substantially straight line as it drives the connecting link  165 . 
     Piston  194  drives connecting link  172 , which is pivotally mounted both to bell crank  150  via bell crank pin  154 , and to the outer arm  184  via inner joint  174 . Outer arm  184 —which forms one of the “side links” of yet another watt linkage mechanism—is also pivotally mounted in fixed relation to the engine block via outermost bearing  185 . The bell crank  150  forms the other “side link” of this watt linkage mechanism. Because of this watt linkage mechanism, the connection end of piston  194  is able to travel in a substantially straight line as it drives the connecting link  172 . 
     Piston  196  drives connecting link  175 , which is pivotally mounted both to bell crank  150  via bell crank pin  156 , and to outer arm  186  via inner joint  179 . Outer arm  186 —which forms one of the “side links” of yet one more watt linkage mechanism—is also pivotally mounted in fixed relation to the engine block via outermost bearing  187 . The bell crank  150  forms the other “side link” of this watt linkage mechanism. Because of this watt linkage mechanism, the connection end of piston  196  is able to travel in a substantially straight line as it drives the connecting link  175 . 
     Because pistons  190 ,  192 ,  194 , and  196  are able to travel in substantially straight line paths, this practically eliminates the side loads exerted by the pistons on their respective cylinders, and minimizes if not eliminates the need for a piston skirt. 
     It is worth noting that the distance between the outermost bearing  181  and the inner joint  164  is equal to the distance between fulcrum  142  and bell crank pin  144 . Likewise, the distance between outermost bearing  183  and inner joint  169  is equal to the distance between fulcrum  142  and bell crank pin  146 . Similarly, the distance between outermost bearing  185  and inner joint  174  is equal to the distance between fulcrum  152  and bell crank pin  154 . Finally, the distance between outermost bearing  187  and inner joint  179  is equal to the distance between fulcrum  152  and bell crank pin  156 . 
     In  FIG. 1 , bell crank pin  144  is radially disposed, with respect to fulcrum  142 , approximately 180 degrees from the bell crank pin  146 . Likewise bell crank pin  154  is radially disposed, with respect to fulcrum  152 , approximately 180 degrees from bell crank pin  156 . With this arrangement, piston  190  may be oriented in parallel with piston  192 , and piston  194  may be oriented in parallel with piston  196 . 
     The angle between the line intersecting bell crank pin  144  and fulcrum  142  and the line connecting fulcrum  142  with bell crank pin  132  is acute, that is, less then 90 degrees. The angle between the line connecting bell crank  150  with bell crank pin  154  and the line connecting fulcrum  152  with bell crank pin  134  is obtuse, that is more then 90 degrees. Preferably, values for these acute and obtuse angles are chosen so that the piston  192  can be collinearly oriented with piston  196  and so that piston  190  can be collinearly oriented with piston  194 . In such an arrangement, pistons  190  and  194  either move towards each other at the same time or away from each other at the same time. Likewise, pistons  192  and  196  move towards each other or away from each other at the same time. 
     With this symmetrical arrangement of the pistons, the forces exerted by the pistons  190 ,  192 ,  194 , and  196  are mostly balanced through the force transfer mechanism itself and not through the crankshaft  110 . 
       FIG. 4  is an orthogonal view of a more detailed embodiment of the force transfer mechanism  100  in  FIG. 1 . This embodiment employs opposing pairs of bell cranks and connecting links on either side of the plane in which the four pistons  190 ,  192 ,  194 , and  196  travel. This eliminates moments perpendicular to the travel of the pistons  190 ,  192 ,  194 , and  196  that the force of the pistons on the piston pins  162 ,  167 ,  172 , or  177  would otherwise generate. Accordingly, a bell crank  240  is disposed opposite bell crank  140  and another bell crank  250  is disposed opposite bell crank  150 . Connecting link  130 , which comprises link plates  230  and  231 , is also matched by another connecting link comprising links  232  and  233 . (See also  FIG. 5 ). Connecting link  165 , which comprises link plates  265  and  266 , are also matched with another connecting link comprising link plates  267  and  268 . Connecting link  175  comprising link plates  275  and  276  are also matched by an opposing connecting link comprising link plates  277  and  278 . Although not illustrated in  FIG. 4 , a similar multiple-link-plate configuration would be supplied for connecting links  160  and  170 . 
       FIG. 5A  depicts a cross-sectional view of the connecting link  130  and its four link plates  230 ,  231 ,  232 , and  233 . The link plates  230 - 233  include bosses  234 . Steel bushings  138  and  148  hold the bosses  234  together. Bronze bushings  139  and  149 , which are lubricated, ride on the steel bushings  138  and  148 , enabling the journal sleeves  141 ,  151 ,  241 , and  251  of the bell cranks  140 ,  150 ,  240 , and  250  to pivot about the bosses  234 . 
       FIG. 5B  depicts a cross-sectional view of the connecting link  165  and its four link plates  265 - 268 . Here, for purposes of facilitating engine assembly, it is preferred that in the region of the inner joint  169 , the outer link plates  265  and  268  include bosses  271  that extend all the way through openings  272  in inner link plates  266  and  267 . The cross-sectional views of the other three connecting links  160 ,  170 , and  175 , not shown, are identical. 
       FIGS. 6A ,  6 B, and  6 C are enlarged, orthogonal, and perspective views of the outer arm  180  of  FIG. 1 . Because the outer arms  180 ,  182 ,  184 , and  186  reach inside of the cylinders  197  when the pistons are at their minimum volume, the outer arms have bends  198  in them from the side view and arches  199  across their axes to allow the arms to clear the edge of the cylinder  197 . The outer arms are of sufficient width to provide adequate strength and rigidity in stabilizing the paths of the pistons while still being able to fit inside the cylinders. On the outer arm  186  near the crankshaft  110  (see  FIG. 2 ), the connecting rod may pass between the opening  195  in the arm (see  FIG. 6C ) near the bearings  187 . 
       FIGS. 7-9  are schematic representations of another embodiment of a force transfer mechanism at different stages of piston travel. Force transfer mechanism  500  comprises a bell crank  540  that pivots about a frame-anchored fulcrum  542  and another bell crank  550  that pivots about a frame-anchored fulcrum  552 . A first watt linkage formed between fulcrums  542  and  552  drives a connecting rod  520  to turn a crankshaft (not shown). 
     The crankshaft pin  522  of connecting rod  520 —which is typically mounted to the rod journal (not shown) of the crank shaft rod—moves along circular travel path  526 , thereby turning the crankshaft. The opposite end of the connecting rod  520  pivots about joint  536  of the watt linkage mechanism formed between fulcrums  542  and  552 . Joint  536  travels along the substantially straight travel path  509 . 
     Pistons  590  and  592  drive the bell crank  540  through second and third watt mechanisms. Likewise, pistons  595  and  596  drive the bell crank  550  through fourth and fifth watt linkage mechanisms. 
     Piston rod (or link)  501  of piston  590  is pivotally attached to piston pin  562  of the second watt linkage mechanism formed between the fulcrum  542  and the frame-anchored pivot  581 . The piston pin  562  travels along travel path  505 , which is substantially straight and parallel to the piston rod  501 . The substantially straight travel path of piston rod  501  eliminates or minimizes the necessity of a piston skirt that would slide with piston  590  along the walls of cylinder  591 . 
     Piston rod (or link)  502  of piston  592  is pivotally attached to piston pin  567  of the third watt linkage mechanism formed between the fulcrum  542  and the frame-anchored pivot  583 . The piston pin  567  travels along travel path  506 , which is substantially straight and parallel to the piston rod  502 . The substantially straight travel path of piston rod  502  eliminates or minimizes the necessity of a piston skirt that would slide with piston  592  along the walls of cylinder  593 . 
     Piston rod  503  (or link) of piston  596  is pivotally attached to piston pin  577  of the fourth watt linkage mechanism formed between the fulcrum  552  and the frame-anchored pivot  584 . The piston pin  577  travels along travel path  507 , which is substantially straight and parallel to the piston rod  503 . The substantially straight travel path of piston rod  503  eliminates or minimizes the necessity of a piston skirt that would slide with piston  596  along the walls of cylinder  597 . 
     Piston rod (or link)  504  of piston  595  is pivotally attached to piston pin  572  of the fifth watt linkage mechanism formed between the fulcrum  552  and the frame-anchored pivot  585 . The piston pin  572  travels along travel path  508 , which is substantially straight and parallel to the piston rod  504 . The substantially straight travel path of piston rod  504  eliminates or minimizes the necessity of a piston skirt that would slide with piston  595  along the walls of cylinder  595 . 
       FIG. 10  is a schematic representation of an alternative embodiment of the force transfer mechanism of the present invention. In  FIG. 10 , as in  FIGS. 7-9  the travel paths  505 - 508  are coplanar and substantially parallel with each other. But unlike the embodiment shown in  FIGS. 7-9 , travel path  506  is not substantially collinear with travel path  507  and travel path  505  is not substantially collinear with travel path  508 . Rather, in  FIG. 10 , the relative spacing and arrangement of fulcrums  542  and  552  is such that travel path  509  is substantially perpendicular to travel paths  505 - 508 . 
       FIG. 11  is a dimensional diagram of a watt linkage mechanism usable in the present invention. A dynamic and static force analysis can be used to predict approximately optimal ratios between the length a of side links  710 , length b of the coupler link  720 , and the distance f between frame-fixed pivots  725  and  730  for each of the 4 piston-driven watt linkages—where substantially straight travel of the piston rods is highly beneficial—and for the connecting-rod-driving watt linkage, where producing a substantially straight path is not as important. Presently, these ratios are believed to be approximately as shown in the following table: 
     
       
         
               
             
               
               
               
               
             
               
               
               
               
             
           
               
                 TABLE 1 
               
             
             
               
                   
               
               
                 Preferred dimensional ratios between watt linkage members 
               
             
          
           
               
                   
                 Ratio 
                 Piston-driven watt linkages 
                 CR-driving watt linkage 
               
               
                   
                   
               
             
          
           
               
                   
                 a/b 
                 1.6 
                 2 
               
               
                   
                 a/f 
                 0.507 
                 0.516 
               
               
                   
                   
               
             
          
         
       
     
     The absolute values of a, b, and f may vary depending on the size and power of the engine one wishes to make. 
     Once the values for a, b, and f are determined, it is possible, using basic trigonometry, to determine the angle D between the line intersecting frame-fixed pivots  725  and  730  and the line intersecting fixed pivot  725  and moving pivot  745  when the midpoint  750  (which travels along travel path  755 ) of coupler link  720  intersects the line connecting frame-fixed pivots  725  and  730 . Then, if one treats the frame-fixed pivot as the center of a hypothetical x-y coordinate system, with the x-axis being collinear with the line intersecting frame-fixed pivot  725  and moving pivot  745  when the midpoint  750  intersects the line connecting frame-fixed pivots  725  and  730 , then it is also possible to determine the x- and y-axes displacements e and d, respectively, of pivot  730  with respect to pivot  725 . These values can then be used to determine the relative spacing and angular orientations of the fixed-frame pivots  542 ,  552  and  581 - 584  ( FIG. 7 ) of an engine built to simulate the embodiment depicted in  FIG. 7 . 
     Basic trigonometry can also be used to determine the angular travel of each of the pivots  725 ,  740 ,  745 , and  730  of the watt linkage through the travel path  755  of the piston. The necessary diameters of each of the pivots will be a function of numerous variables, including the strength properties of the materials selected and the moments and forces each pivot needs to be able to withstand in operation. 
     Based on some preliminary modeling and analysis, I believe that a four-cylinder engine built in accordance with my H-cross configured force-transfer mechanism could deliver power equal to that of a standard in-line four-cylinder engine with as little as 28% of the oil shear area times shear distance that one would find in a comparable in-line four-cylinder engine. The following tables compare the expected shear area times distance traveled between the two engines, using reasonable assumptions for the bearing diameters: 
     
       
         
               
             
               
               
               
               
               
             
               
               
               
               
               
             
           
               
                 TABLE 2 
               
             
             
               
                   
               
               
                 anticipated bearing dimensions for H-cross four-cylinder engine 
               
             
          
           
               
                   
                 Number 
                 Degrees of rota- 
                 bearing 
                 bearing 
               
               
                 Bearing 
                 of 
                 tion/revolution 
                 diameter 
                 length 
               
               
                 type 
                 bearings 
                 of crankshaft 
                 (in.) 
                 (in.) 
               
               
                   
               
             
          
           
               
                 Outermost bearing 
                 8 
                 120 
                 .5 
                 .75 
               
               
                 Inner bearing of 
                 8 
                 118 
                 .5 
                 .75 
               
               
                 outer link 
               
               
                 piston pins 
                 4 
                 60 
                 .875 
                 1 
               
               
                 bell crank pins 
                 12 
                 118 
                 .75 
                 .75 
               
               
                 bell crank mains 
                 4 
                 120 
                 1.25 
                 .75 
               
               
                 Connecting rod 
                 1 
                 60 
                 1 
                 1 
               
               
                 wrist pin 
               
               
                 rod journal 
                 1 
                 360 
                 1.875 
                 1 
               
               
                 bearing 
               
               
                 Crankshaft main 
                 2 
                 360 
                 2 
                 1 
               
               
                 bearing 
               
               
                   
               
             
          
         
       
     
     
       
         
               
             
               
               
               
               
             
               
               
               
               
             
           
               
                 TABLE 3 
               
             
             
               
                   
               
               
                 oil shear-area times shear distance traveled 
               
               
                 for an H-cross four-cylinder 
               
             
          
           
               
                   
                 oil shear 
                 shear area travel 
                 product of shear 
               
               
                   
                 area/ 
                 distance per 
                 area times shear 
               
               
                 Bearing 
                 bearing 
                 revolution of 
                 travel times number 
               
               
                 type 
                 (in.{circumflex over ( )}2) 
                 crankshaft (in.) 
                 of bearings (in.{circumflex over ( )}3) 
               
               
                   
               
             
          
           
               
                 outermost bearing 
                 2.36 
                 0.524 
                 4.93 
               
               
                 inner bearing of 
                 2.36 
                 0.515 
                 4.85 
               
               
                 outer link 
               
               
                 piston pins 
                 2.75 
                 0.458 
                 5.04 
               
               
                 bell crank pins 
                 3.53 
                 0.772 
                 16.4 
               
               
                 bell crank mains 
                 5.89 
                 1.31 
                 15.4 
               
               
                 connecting rod 
                 3.14 
                 0.524 
                 1.64 
               
               
                 wrist pin 
               
               
                 rod journal 
                 5.89 
                 5.89 
                 34.7 
               
               
                 bearing 
               
               
                 crankshaft main 
                 6.28 
                 6.28 
                 79.0 
               
               
                 bearing 
                   
                   
                   
               
               
                 TOTAL 
                 — 
                 — 
                 162 
               
               
                   
               
             
          
         
       
     
     
       
         
               
             
               
               
               
               
               
             
               
               
               
               
               
             
           
               
                 TABLE 4 
               
             
             
               
                   
               
               
                 comparable dimensions for a standard in-line 4 cylinder engine 
               
             
          
           
               
                   
                 Number 
                 Degrees of rota- 
                 bearing 
                 bearing 
               
               
                 Bearing 
                 of 
                 tion/revolution 
                 diameter 
                 length 
               
               
                 type 
                 bearings 
                 of crankshaft 
                 (in.) 
                 (in.) 
               
               
                   
               
             
          
           
               
                 rod journal bearing 
                 4 
                 360 
                 2 
                 1 
               
               
                 main bearing 
                 5 
                 360 
                 2 
                 1 
               
               
                 wrist pin 
                 4 
                 60 
                 1 
                 1 
               
               
                 piston skirts 
                 4 
                 n/a 
                 n/a 
                 n/a 
               
               
                   
               
             
          
         
       
     
     
       
         
               
             
               
               
               
               
             
               
               
               
               
             
           
               
                 TABLE 5 
               
             
             
               
                   
               
               
                 oil shear-area times shear distance traveled 
               
               
                 for a standard in-line 4 cylinder engine 
               
             
          
           
               
                   
                 oil shear 
                 shear area 
                 product of shear 
               
               
                   
                 area/ 
                 travel distance 
                 area times shear 
               
               
                 Bearing 
                 bearing 
                 per revolution of 
                 travel times number 
               
               
                 type 
                 (in.{circumflex over ( )}2) 
                 crankshaft (in.) 
                 of bearings (in.{circumflex over ( )}3) 
               
               
                   
               
             
          
           
               
                 rod journal 
                 6.28 
                 6.28 
                 158 
               
               
                 bearing 
               
               
                 main bearing 
                 6.28 
                 6.28 
                 197 
               
               
                 wrist pin 
                 3.14 
                 3.14 
                 39.5 
               
               
                 piston skirts 
                 8 
                 6 
                 192 
               
               
                 TOTAL 
                 — 
                 — 
                 587 
               
               
                   
               
             
          
         
       
     
     The bearing diameter assumptions for the standard in-line 4-cylinder engine are relatively close to the actual diameters seen in modern-day care engines. The bearing diameter assumptions for the inventive embodiment, although somewhat smaller, are reasonable based on the anticipated loading on each of the bearings and because many of the bearings will have a diameter-to-length ratio that is relatively close to one. 
     The value of reducing the total shear area times distance traveled per crankshaft revolution is confirmed by the fact that the force needed to overcome mechanical friction between the lubricated surfaces of the bearing is not only directly proportional to the surface area of the sliding surfaces, but also directly proportional to the relative velocity of the sliding surfaces. This relationship is expressed by the following formula:
 
 F=μ*V*A/Y  
 
where μ is the coefficient of viscosity, V is the velocity, A is the sliding surface area, and Y is the distance separating the two sliding surfaces. For purposes of the embodiments disclosed herein, the velocity V is directly proportional to the shear distance traveled per crankshaft revolution times the revolutions per minute (RPM) of the crankshaft.
 
     Because the preferred embodiment reduces the total shear area times distance traveled per crankshaft revolution, the preferred embodiment should reduce the frictional losses of the engine. The invention&#39;s enablement of reductions in the weight and size of the crankshaft contributes to further efficiency gains. Accordingly, it is believed that the resulting reduction in frictional losses will significantly improve the efficiency of the engine, particularly under part-load conditions. 
     It is stressed that the invention is not limited to a 4-cylinder configuration.  FIG. 12  illustrates an embodiment of an eight cylinder configuration with 2 force transfer mechanisms  602  and  604  disposed on opposite sides of the crank shaft (not shown).  FIG. 13  illustrates an embodiment of a 12-cylinder configuration with 3 force transfer mechanisms  802 ,  804  and  806  stacked next to each other on the same side of crankshaft  810 . In this configuration the crankshaft  810  would have four main journals  850 ,  852 ,  854 , and  856  and three rod journals  820 ,  830 , and  840  that are offset 120 degrees apart from each other. 
     It is also stressed that the invention, unless so specified in the claims, is not necessarily limited to a multiple watt-linkage mechanism.  FIG. 14  is an alternative embodiment of a force transfer mechanism according to the present invention which employs a standard crank and slider mechanism to drive the bell cranks. In this embodiment the force transfer mechanism  900  includes only one watt linkage mechanism driven by bell cranks  910  and  920 . Each of the bell cranks  910  and  920  are driven by pistons  940 ,  941 ,  942 , and  943  that are attached via wrist pins  962 ,  963 ,  964 , and  965  to piston rods  950 ,  951 ,  952 , and  953 . In this embodiment the piston rods  950 - 953  do not travel in as straight a line as they would were they driving watt linkage mechanisms. However, their paths are still straighter then what one would find in a standard in-line four-cylinder engine. Accordingly, the lateral loads experienced by each piston  940 - 943  on its corresponding cylinder are still substantially reduced, which reduces the needed length of the piston skirt  930 ,  932 ,  934 , or  936 . 
       FIG. 15  is another alternative embodiment of a force transfer mechanism  1000  that utilizes only a single watt linkage mechanism and uses curved piston cylinders  1010 ,  1020 ,  1030 , and  1040 , to drive the bell cranks  1050  and  1060 . The curved construction of the cylinders  1010 ,  1020 ,  1030 ,  1040 , allows the corresponding pistons  1015 ,  1025 ,  1035 , and  1045 , to travel along substantially curvilinear paths as they drive the bell cranks  1050  and  1060 , thus again eliminating the need for piston skirts. 
       FIG. 16  is an alternative embodiment of a force transfer mechanism  1600  that utilizes four watt linkage mechanisms  1610 ,  1620 ,  1630 , and  1640  to drive bell cranks  1650  and  1660 , which drive two separate two connecting rods  1670  and  1680 . In this embodiment, the pistons  1605 ,  1615 ,  1625 , and  1635  are all be oriented in the same direction, and there is no watt linkage mechanism between the bell cranks  1650  and  1660  and the crankshaft  1690 . In one sub-embodiment, the connecting rods  1670  and  1680  drive two separate crank throws (not shown) on the crankshaft  1690 . In another sub-embodiment, shown in  FIG. 17 , the connecting rods  1670  and  1680  are linked together with a pin  1710 , so that the between-cylinder forces are transferred from one connecting rod to the other. 
       FIG. 18  is yet another alternative embodiment of a force transfer mechanism  1800  that is similar to the force transfer mechanism  1600  of  FIG. 16 , but which utilizes no watt linkages. In this embodiment, pistons  1805 ,  1815 ,  1825 , and  1835  drive bell cranks  1850  and  1860  in the same fashion described and depicted with respect to  FIG. 14 . The force transfer mechanism  1800  still distributes the between-cylinder forces better than the conventional 4-cylinder in-line engine. It also still reduces the necessary length of the piston skirts and the projected shear area times distance traveled per crankshaft revolution. 
     As used in this specification and claims, the term “bell crank” refers broadly to a first-class lever that rotates about a fulcrum and is used to convert the direction of reciprocating movement. The bell crank may include one or more input force points where force is applied to the bell crank, and one or more output force points where the input force is transferred to one or more other bodies. The same point may serve as either an input force point or an output force point, depending on the particular stroke (intake, compression, power, exhaust) that a given piston is in. 
     In a bell crank, the imaginary line connecting the fulcrum to the point on the bell crank where an input force is applied is at an angle, other than 0 or 180 degrees, from the imaginary line connecting the fulcrum to at least one other point on the bell crank where force is transferred to another body. 
     A bell crank, however, may include two or more output force points, one of which is disposed 180 degrees from the input force point, provided that at least one of the other output force points is disposed at some angle other than 0 or 180 degrees from the input force point. For example,  FIGS. 1-3  and  7 - 10  depict bell cranks with three bell crank pins or pivots, the first two of which are disposed approximately 180 degrees (with respect to the frame-fixed fulcrum) from each other, and a third of which is disposed approximately 75 degrees and 115 degrees, respectively, from the first two bell crank pins or pivots. 
     As used in this specification and claims, a bell crank is not limited to cranks that change motion around a 90 degree angle. Nor is it limited to a two-armed lever that shares a fulcrum at the point where the arms join. As used in this specification, even a wheel—in which a force input point is disposed at an angle, with respect to the fulcrum, from a force output point—could serve as a bell crank. Indeed, given that it is preferable that a bell crank be structurally designed to achieve a high strength-to-weight ratio, an optimal bell crank for use as part of an engine force transfer mechanism will be more structurally complex than a simple two- or three-armed lever. 
     As used in this specification and claims, a “watt linkage” refers to what mechanical engineers would schematically characterize as a type of a planar “four-bar linkage” comprising four rigid bodies (one of which is the frame), each of which are attached to two of the other bodies by single joints or pivots to form a closed loop. A watt linkage is further characterized as comprising two “side links” of approximately equal and constant functional length hinged via frame-fixed pivots to a frame and a “coupler link” opposite the frame pivotally connected on either end (via pivots that are not fixed with respect to the frame) to the distal ends of the two “side links.” (The “functional length” of the side link would be the distance between the frame-fixed pivot and the coupler-link pivot of the side link.) The side links and coupler link can rotate to some extent about their pivots, but they are not free to translate with respect to those pivots. As used in this specification and claims, a “watt linkage” does not suggest that any of its four rigid bodies take the form of rods or bars or linear members. 
     It should be understood that when the claims recite a “watt linkage” mechanism, that mechanism may include, as one or more of its four rigid bodies, elements that have already been recited in the claim. For example, this patent application&#39;s originally-filed claim  1  (which may differ from the issued claim  1  that appears in this patent) recites a “first watt linkage comprising a connecting link pivotally connected on opposite ends to first and second bell cranks, thereby mechanically coupling the first bell crank with the second bell crank.” Here, the first and second bell cranks—which have already been recited in the claim—comprise what are, in effect, the “side links” of the “first watt linkage.” 
     As used in the claims, and unless otherwise qualified, the “midpoint of the connecting link”—in the context of a watt linkage mechanism—refers to the functional midpoint of the connecting link, which is the midpoint between the two pivots of the connecting link that join the connecting link with the “side links” of the watt linkage. In theory, the functional midpoint could differ from the actual midpoint if one end of the connecting link extended further beyond its pivot than the connecting link&#39;s opposite end extended beyond its opposite pivot. Unless further qualified, the claim language “midpoint of the connecting link” refers to the functional midpoint, not necessarily the actual midpoint, of the connecting link. 
     As used in this specification and claims, an “H-cross configuration” is a configuration in which 4 piston cylinders are longitudinally oriented in approximately the same plane and in substantially parallel relation to each other, including two opposing pairs of cylinders on opposite sides of a watt-linkage mechanism that transfers the piston forces to a connecting rod. In a “collinear H-cross configuration,” each of the cylinders at the top of the “H” are substantially collinear with a corresponding cylinder at the bottom of the “H.” Examples of “collinear H-cross configurations” are depicted in  FIGS. 1-3  and  7 - 9 . In an “offset H-cross configuration,” the cylinders at the top of the “H” are offset from (but still on the same side of the crankshaft as) the cylinders at the bottom of the “H.” An example of an “offset H-cross configuration” is shown in  FIG. 10 . Therefore, it should be understood that the phrase “H-cross configuration,” unless otherwise qualified, encompasses both collinear and offset H-cross configurations. 
     It will be observed that in the drawings accompanying the specification, each force transfer mechanism drives a corresponding connecting rod along a trajectory which—although not entirely linear—is substantially perpendicular to the substantially parallel trajectories of the reciprocating pistons. The substantially parallel trajectories of the reciprocating pistons (more specifically, of the piston pins at the end of the piston rods) are illustrated, for example, in  FIGS. 7-10  by reference numbers  505 - 508 . The non-linear trajectory of the connecting rod is illustrated best in  FIGS. 1-3  and  14 . In all of the illustrated embodiments, the connecting rod is oriented—and its small end travels along a trajectory that is oriented—more perpendicular than parallel to the mostly parallel trajectories of the pistons powering the connecting rod through the force transfer mechanism. Accordingly, to the extent that the claims make reference to a trajectory that is “substantially perpendicular” to a reference trajectory (or trajectories), it is intended that the phrase encompass trajectories that are closer to the perpendicular than to the parallel to the reference trajectory (or trajectories). 
     Although the foregoing specific details describe various embodiments of the invention, persons reasonably skilled in the art will recognize that various changes may be made in the details of the apparatus of this invention without departing from the spirit and scope of the invention as defined in the appended claims. 
     The present invention includes several independently meritorious inventive aspects and advantages. Unless compelled by the claim language itself, the claims should not be construed to be limited to structures that incorporate all of the inventive aspects, or enjoy all of the advantages, disclosed herein. 
     It is well established that the claims of the patent serve an important public notice function to potential competitors—enabling them to not only determine what is covered, but also what is not covered—by the patent. And a number of Federal Circuit decisions have emphasized the importance of discerning the patentee&#39;s intent—as expressed in the specification—in construing the claims of the patent. 
     But defendants in patent infringement suits—while arguing the importance of this public notice function—often seek strained and uncharitable constructions of the claims that would render them either nonsensical, too narrow to have any significant value, or so broad that the claim is anticipated by the prior art. Defendants are apt to mercilessly flog minor grammatical, typographical, or syntactical flaws, if any, in the claims or specification, forgetting that patents are generally written by—and for—engineers and technicians, not by and for grammatical perfectionists and English language PhD&#39;s. Furthermore, defendants frequently misconstrue the specification and prosecution history in claim construction briefs and hearings in an effort to get courts to import all kinds of contrived and novel limitations into the construction of the claims. They also frequently strive to—in essence—rewrite the claims so that they do not cover the accused device. 
     Accordingly, I wish to make my intentions clear—and at the same time put potential competitors on clear public notice. It is my intent that the claims receive a liberal construction and be interpreted to uphold and not destroy the right of the inventor. It is my intent that the claim terms be construed in a charitable and common-sensical manner. It is my intent that the claim terms be construed as broadly as practicable while preserving the validity of the claims. It is my intent that the claim terms be construed in a manner consistent with the context of the overall claim language and the specification, without importing extraneous limitations from the specification or other sources into the claims, and without confining the scope of the claims to the exact representations depicted in the specification or drawings. It is also my intent that not each and every term of the claim be systematically defined and rewritten. Claim terms and phrases should be construed only to the extent that it will provide helpful, clarifying guidance to the jury, or to the extent needed to resolve a legitimate, good faith dispute that is material to the questions of validity or infringement. Otherwise, simple claim terms and phrases should be presented to the jury without any potentially confusing and difficult-to-apply definitional construction. 
     It is also to be understood that the terminology employed in the Summary of the Invention and Detailed Description sections of this application is for the purpose of describing particular embodiments. Unless the context clearly demonstrates otherwise, is not intended to be limiting. In this specification and the appended claims, the singular forms “a,” “tan” and “the” include plural references unless the context clearly dictates otherwise. Conversely, it is contemplated that the claims may be drafted to exclude any optional element or be further limited using exclusive terminology as “solely,” “only” and the like in connection with the recitation of claim elements or by use of a “negative” limitation. It is also contemplated that any optional feature of the inventive variations described herein may be set forth and claimed independently, or in combination with any one or more of the features described herein. 
     The headquarters building of the World Intellectual Property Organization bears the following inscription: “Human genius is the source of all works of art and invention; these works are the guarantee of a life worthy of me; it is the duty of the State to ensure with diligence the protection of the arts and inventions.” It is my intent that the claims of this patent be construed—and ultimately enforced, if necessary—in a manner worthy of this mandate.