Abstract:
The present invention is directed to improved designs and methods for improving engine fuel efficiency by providing two-stage engine variable compression in running engines using connecting rod force reversals to rotate eccentric bushings to change the connecting rod length. Compression ratio changes are initiated by shifting a block-mounted cam such that it engages and flips a bi-stable toggle on the connecting rod. The clutch mechanism latches the eccentric at the eccentric rotation end point, whereupon the connecting rod acts as a rigid rod. The invention includes novel configurations of the lubricated journal bearing between the connecting rod and the eccentric that modify the squeeze film bearing effects and resulting friction. These configurations reduce the peak eccentric torque carried by the clutch mechanism while facilitating eccentric rotation at lower torque.

Description:
FIELD OF THE INVENTION 
       [0001]    The present invention is directed to improving engine fuel efficiency by providing adjustable length internal combustion engine connecting rods that link the engine crankshaft to the pistons, wherein the rod length may be changed in a running engine to raise or lower the compression ratio as a function of engine operating conditions. More specifically, the invention is directed to two-stage adjustable length connecting rods in which the adjustable element is an eccentric connection of the rod end to the piston wrist pin that is actuated by the periodically reversing gas and inertial dynamic forces acting on the rod and eccentric assembly, and latched in either of two positions resulting in two different effective rod lengths and corresponding compression ratios. 
       BACKGROUND OF THE INVENTION 
       [0002]    The present invention comprises improvements to generally known two-stage adjustable length connecting rods in which the adjustable element is an eccentric connection of the rod end to the piston pin that is actuated by the periodically reversing gas and inertial dynamic forces acting on the rod and eccentric assembly, and latched in either of two positions resulting in two different effective rod lengths and corresponding compression ratios. U.S. Pat. No. 7,685,974 issued to Berger provides extensive background information including prior art, utility of two-stage variable compression ratio for improved engine efficiency, and nature of the periodic gas and inertial dynamic forces operating on the rod and eccentric assembly. &#39;974 is incorporated by reference, and a brief summary follows. The discussion is directed to spark ignition throttle controlled internal combustion engines typically used in light duty vehicles, although it is obvious that the invention is applicable to compression ignition engines and other reciprocating machines such as compressors. 
         [0003]    Internal combustion engine compression ratio is defined as the volume between the piston and the cylinder head at bottom dead center divided by the volume at top dead center. In typical prior art engines the compression ratio is fixed by the engine design and is not an adjustable operating variable. In high compression engines the volume at top dead center is relatively small, and the compressed fuel-air mixture is at a higher temperature and pressure than in low compression engines at the moment of ignition. Increasing temperature and pressure increase the engine thermodynamic efficiency by increasing the combustion pressure that drives the piston toward bottom dead center on the power stroke, thereby producing more useful work from a given quantity of fuel. The temperature and pressure must, however, remain below the knock-limit. Below the knock limit, the fuel-air mixture combusts as a progressive flame front passing through the combustion chamber volume, and results in a smooth pressure rise to drive the piston. Above the knock limit, a portion of the fuel-air mixture in the combustion chamber undergoes very rapid bulk combustion that produces a pressure spike resulting in audible noise and potential engine damage. The knock-limit is affected by a number of variables other than compression ratio including fuel chemical composition, inlet manifold pressure, engine temperature, and ignition timing relative to piston position. Modern engines typically have knock sensors that detect incipient knock. These allow the engine control module to automatically make adjustments, e.g. varying ignition timing, to keep the engine operating at the highest possible efficiency without knock over the full engine load and speed range, while compensating for other variables such as fuel chemistry and engine temperature. Late or retarded ignition timing from the maximum torque value reduces knock, but also reduces engine efficiency and increases heat transfer losses. 
         [0004]    The fixed compression ratio of prior art engines is a compromise that balances maximum output torque against light load efficiency. Maximum output torque at a given speed occurs when the throttle is fully open, thereby providing the maximum gas pressure in the inlet system. This gas may be a fuel-air mixture in manifold or port injection systems, or air alone in direct cylinder injection systems in which fuel is added to the air in the cylinder. In a naturally aspirated engine this pressure is slightly below atmospheric pressure because of flow losses, and in a turbocharger or supercharger pressure boosted engine it is above atmospheric pressure. The cylinder at the beginning of the compression stroke is filled with a charge of gas slightly below the inlet manifold pressure, since inlet system flow losses prevent complete pressure equilibration between the inlet system and the cylinder. The compression ratio is chosen so that at the end of the compression stroke the gas charge is at a temperature and pressure that results in maximum output torque with optimized ignition timing and without knock. A pressure boosted engine generally requires a lower compression ratio to avoid knock at maximum torque since the initial gas pressure and charge mass is higher. At light load the throttle reduces the inlet system pressure and the initial cylinder pressure resulting in a smaller charge mass. This leads to lower than optimum gas charge temperature and pressure at the end of the compression stroke, and consequent reduced efficiency. Since the typical duty cycle of light duty vehicles consists of predominantly light load operation, the non-optimum compression ratio results in significant efficiency losses. 
         [0005]    It is generally known that that a capability to adjust compression ratio in a running engine as a function of operating conditions has potential to improve both light load efficiency and maximum output torque compared to a fixed compression ratio. High compression at light loads compresses the low mass light load gas charge into a reduced combustion chamber volume. This increases the gas charge temperature and pressure at the end of the compression stroke to a more optimum level and increases the thermodynamic efficiency. Reduced compression at maximum output torque compresses the maximum load gas charge into an increased combustion chamber volume. This allows a larger charge mass without exceeding the optimum temperature and pressure at the end of the compression stroke, allowing higher output torque. This is a particular advantage in pressure boosted engines since it permits a smaller displacement engine to provide the required maximum output torque, offering potential engine size and weight savings. 
         [0006]    Various approaches to adjusting the compression ratio are known, and are summarized in &#39;974. Two-stage adjustable length connecting rods are the subject of &#39;974 and a number of other patents, and are attractive because they require minimum changes to the engine and add little overall size or weight. This approach employs two-stage adjustment in which the rod may be switched between two fixed lengths, since it is simpler than continuous adjustment and provides nearly the same benefit. As in a conventional fixed compression ratio engine, the crankshaft axis is fixed relative to the cylinder head. An increase in connecting rod length reduces the volume at top dead center and increases the compression ratio, while a decrease in length increases the volume at top dead center and reduces the compression ratio. Many mechanisms are disclosed in the prior art to carry out the adjustment, and a number use an eccentric bushing between the piston wristpin and the small end of the connecting rod that is rotated relative to the connecting rod in a journal bearing to adjust the effective connecting rod length. Although there are exceptions, typical prior art mechanisms are self-powered in that the alternating compressive and tensile forces on the connecting rod and eccentric bushing during the engine cycle generate torque that rotates the eccentric to change the effective rod length. Mechanical stops are incorporated to limit the eccentric bushing rotation to a specific angle. A bi-stable mechanical latch mounted on the moving connecting rod locks the eccentric into either the high compression or the low compression position, and is reset by a stationary trigger mechanism that interacts with the latch on the moving connecting rod to change the compression ratio. When the latch is reset it is biased to disengage the eccentric, and then engage it again after the alternating rod forces rotate the eccentric to the new selected position. The latch typically disengages only under low load as the alternating rod load reverses, and the eccentric will only move toward the new selected position when the alternating rod load is in the correct direction. Compressive rod force is required to rotate the eccentric within the journal bearing to reduce the effective rod length, and tensile rod force is required to rotate the eccentric in the opposite direction to increase the effective rod length. The latch reengages the eccentric only when the eccentric reaches the new position. One or more engine revolutions after latch reset are required to obtain the connecting rod force variations needed to release the latch and rotate the eccentric to the new position so that the latch may reengage. 
         [0007]    Three principal approaches are proposed in the prior art to allow a stationary trigger mechanism to reset the latch on the moving connecting rod to change the compression ratio: hydrostatic fluid interaction, hydrodynamic fluid interaction, and mechanical cam interaction. Hydrostatic fluid interaction employs passages in the engine block, crankshaft and connecting rod such that controllable oil pressure may be transmitted from a stationary source through the rotating interfaces to a hydraulic plunger that shifts the latch mechanism between positions on the moving connecting rod. Hydrodynamic fluid interaction employs stationary oil nozzles mounted to the engine block that direct momentary oil jets to exert forces on control surfaces carried by the moving connecting rod that move to shift the latch between positions. Mechanical cam interaction employs cam surfaces mounted to the engine block that may be controllably shifted to contact and shift the latch between positions on the moving connecting rod. Hybrid approaches, e.g. mechanical cam surfaces that shift valves in the connecting rod that redirect pressurized oil to hydrostatically shift the latches between positions, are also known. 
         [0008]    The two-stage adjustable length connecting rods of &#39;974 and related prior art have an inherent indeterminacy related to the fact that alternating compressive and tensile forces on the connecting rod and eccentric bushing during the engine cycle are used to rotate the eccentric to change the effective rod length. Ideally, one or two engine revolutions after latch reset are required to obtain the connecting rod force variations needed to release the latch and rotate the eccentric to the new position so that the latch may reengage. If, however, the eccentric does not rotate enough to reach the new position and engage the latch while the rod force is in the correct direction, the subsequent force reversal will reverse the initial eccentric rotation direction. This causes a loss of the desired rotation, and may result in situations in which the eccentric rotational position cycles with each rod force direction, and full rotation and latch engagement is either slow or never completed. Such unproductive cycling may also cause wear and noise. Since maximum compressive rod forces include compression and power stroke gas loading and are typically much larger than primarily inertial tensile rod forces, it may be more difficult to lengthen the rod and increase the compression ratio. 
         [0009]    The relationship between the eccentric geometry and the coefficient of friction in the journal bearing between the eccentric and the connecting rod is critical in achieving reliable eccentric rotation while minimizing stress on the latch. The parameters are shown in  FIG. 12 : 
         [0010]    R is the outside radius of the eccentric  108  which rotates within the journal formed by the small end  109  of the connecting rod body  110 ; 
         [0011]    r is the eccentric offset between the journal center of the piston wrist pin  107  and the center of the outside diameter of the eccentric  108 ; 
         [0012]    α is the rotational angle between a line  1200  perpendicular to the centerline  1201  of the connecting rod body  110 ; 
         [0013]    F is the instantaneous value of the tensile or compressive force on the connecting rod  100 ; 
         [0014]    μ is the coefficient of friction between the eccentric  108  rotating within the small end  109  of the connecting rod body  110 ; 
         [0015]    M f  is the torque on the eccentric  108  generated by the friction between the eccentric  108  and the small end  109  of the connecting rod body  110 ; and 
         [0016]    M r  is the reaction torque on the eccentric  108  generated by the force F acting on the eccentric offset. 
         [0000]    
       
      
       M 
       f 
       =FμR  
      
     
         [0000]        M   r   =Fr  cos α
 
         [0017]    If the reaction torque M r  is greater than the friction torque M f  the eccentric will rotate in the direction of the reaction torque, and if it is less it will not rotate. It is useful to define a parameter Z as the ratio of the reaction torque to the friction torque: 
         [0000]        Z=M   r   /M   f   =Fr  cos α/ FμR=r  cos α/μ R  
 
         [0018]    If Z is greater than 1 rotation will take place, and if it is less than 1 it will not. A large Z therefore provides higher assurance of successful rotation between latched positions. A smaller Z however, counteracts a larger portion of the reaction torque with the friction torque and reduces the load on the latch mechanism, while still permitting rotation so long as Z has a value greater than 1. The following example is based on a 75 mm diameter bore and piston: 
         [0019]    Maximum axial force F 25,700 N, 
         [0020]    Rod length adjustment range 4 mm, 
         [0021]    Eccentric offset r 3.5 mm, 
         [0022]    Eccentric rotational angle α+/−55 degrees, 
         [0023]    Eccentric outside radius R 17 mm, and 
         [0024]    Friction coefficient μ 0.05. 
         [0025]    Reaction torque, M r =Fr cos α=51.6 Nm 
         [0026]    Friction torque, M f =FμR=21.8 Nm 
         [0027]    Net rotational torque M n =M r −M f =29.8 Nm 
         [0000]    
       
         
           
             Z 
             = 
             
               
                 
                   M 
                   r 
                 
                 
                   M 
                   f 
                 
               
               = 
               2.37 
             
           
         
       
     
         [0028]    In this example rotation is possible and the net rotational torque M n  that is carried by the latch is reduced by the frictional torque M f . 
         [0029]    Except for the friction coefficient μ, the parameters affecting Z are geometric design parameters. The friction coefficient μ in contrast is only in part a function of the choice of design parameters including the eccentric  108  and rod small end  109  materials, contacting surface finishes, any coatings, and the lubricant formulation. It is also believed to be affected by hydrodynamic effects, particularly transient squeeze film lubrication driven by alternating connecting rod compressive and tensile forces acting on the journal bearing oil film between the connecting rod small end  109  and eccentric bushing  108 . While not wishing to be bound by theory, it is believed that squeeze film lubrication is an important component of piston wrist pin lubrication, and is similarly important in determining the instantaneous friction coefficient μ between the eccentric and the rod, and that it is not adequately addressed in the prior art. 
         [0030]    I. Elsion et al./ Wear  261 (2006) 785-791 describes experimental investigation of piston wrist pin lubrication by rotationally oscillating a piston wrist pin within a journal that is clamped from opposite sides by an applied load. They report a mixed lubrication regime friction coefficient μ in the range of 0.03 to 0.06 for a steel pin and an aluminum journal, and show modest effects of experimental coatings and engineered surface finishes. Since the clamping load is constant rather than periodically reversing, these experiments do not provide information on transient squeeze film lubrication effects in which the bearing is fully hydrodynamic and there is no metal to metal contact. They do, however, provide information on the mixed lubrication friction coefficient when squeeze film lubrication is not occurring. 
         [0031]    In squeeze film lubrication fluid forces momentarily separate two approaching solid surfaces in oil-flooded environments, forming a hydrodynamic bearing with a much lower friction coefficient than mixed lubrication. The hydrodynamic bearing support force F S  for parallel configurations is given by: 
         [0000]    
       
         
           
             
               F 
               S 
             
             = 
             
               C 
                
               
                 
                   
                     μ 
                     l 
                   
                    
                   
                     L 
                     4 
                   
                    
                   V 
                 
                 
                   h 
                   3 
                 
               
             
           
         
       
     
         [0032]    where μ l  is the lubricant viscosity at the operating temperature, L is the shortest flow path length from the center of the bearing area to the edge, V is the perpendicular velocity between the two surfaces, h is the separation between the surfaces, and C is a constant determined by the bearing geometry. It is believed that each time the rod alternates between compressive and tensile force, the journal oil film between the connecting rod small end  101  and eccentric bushing  102  forms a transient low friction hydrodynamic bearing. The time duration of this transient increases with oil viscosity μ l  and reduces with increased force F, resulting in variations with engine speed, load and temperature. 
         [0033]    Transient low friction hydrodynamic bearing effects have the positive effect of facilitating eccentric rotation during rod length change, particularly during the rod length increases to raise compression wherein only relatively low inertial forces are available to provide the rotational torque. Conversely, transient low friction hydrodynamic bearing effects have the negative effect of increasing the torque load on the eccentric latch mechanism from cylinder pressure loading, particularly when maintaining maximum rod length during high compression operation. 
         [0034]    In summary, prior art eccentric bushing variable compression connecting rods may not achieve reliable length shifts under all conditions, and have squeeze film bearing transients in the journal bearing between the connecting rod small end and eccentric bushing that are counterproductive in some operating modes. A need therefore exists for improvements that address these issues. 
       SUMMARY OF THE INVENTION 
       [0035]    The present invention is directed to designs and methods for providing two-stage engine variable compression using eccentric bushings to change connecting rod length that provide reliable and consistent rod length shifts. It is further directed to providing robust mechanisms that minimize volume and weight, incorporate simple controls, and minimize changes to engines. 
         [0036]    According to one aspect of the invention, a switchable one-way clutch mechanism such as ratchet is used as the latch that controls the eccentric rotation rather than the prior art latches that engage only at the eccentric rotation end points. This mechanism captures incremental rotation in the required direction even if the eccentric does not rotate the full travel in a given connecting rod force reversal cycle, so that the full travel may be accumulated over more than one force reversal cycle. The clutch mechanism then latches the eccentric at the eccentric rotation end point. When the clutch direction is switched, the eccentric is free to rotate in the opposite direction with the same incremental rotation capability. This incremental motion capture capability results in more reliable compression ratio shifting by eliminating the possibility of non-productive cycling. Preferably when the clutch direction is switched it maintains its original locking direction until the rod load and resulting torque on the eccentric are at a low level and in the new rotation direction. This pre-set characteristic allows the clutch direction to be switched at any point in the engine cycle, while assuring that the clutch releases only when the force direction and magnitude are favorable. 
         [0037]    A second aspect of the invention modifies the squeeze film interface in the journal bearing between the connecting rod and the eccentric to reduce the peak cylinder loads on the eccentric latching mechanism while facilitating the inertial load driven eccentric rotation from low to high compression. This is accomplished by providing flow passages such as grooves in selected portions of the cylindrical surfaces forming the journal bearing. These flow passages are arranged to reduce the shortest oil flow path L from the center of the squeeze film bearing area to the edge. Since the squeeze bearing force F s  is proportional to 1/L 4 , reductions in L provide effective means of reducing F s  and minimizing squeeze film low friction transients in the oil film in the journal bearing when the applied load is compressing the oil film towards the portions of the bearing having the flow passages. In a preferred embodiment an array of multiple axial grooves are formed in the cylindrical journal bore in the connecting rod small end over the half of its circumference that carries high compressive connecting rod loads transmitted from the eccentric. This has the effect of suppressing the squeeze film bearing effect and assuring that the friction coefficient μ is the higher mixed lubrication friction coefficient and not the lower hydrodynamic coefficient, thus reducing the load on the latch during these high compressive load intervals. Transient hydrodynamic squeeze film lubrication is, however, retained during the lower force tensile load intervals, facilitating eccentric rotation. 
         [0038]    A third aspect of the invention comprises a spring-loaded toggle that may be shifted between two stable positions in response to an engine control signal to pre-set the switchable one-way clutch mechanism. In the first stable position the spring-loaded toggle applies force to the one-way clutch mechanism such that the clutch is biased to allow eccentric rotation that increases the rod length, and in the second stable position the toggle applies force to the one-way clutch mechanism such that the clutch is biased to allow eccentric rotation that decreases the rod length. The toggle force is preferably set to a value insufficient to disengage the clutch when it is highly loaded, preventing abrupt and potentially damaging eccentric rotation and rod length change. Instead the clutch only disengages when the rod force reverses and the eccentric torque passes through zero. After disengagement in one direction, the applied toggle force biases the one-way clutch so that the eccentric rotation can only take place in the opposite direction. These characteristics allow the toggle to be shifted between the two stable positions and thereby pre-set at any point in the engine cycle, and only complete the action when the conditions are favorable. In this disclosure control cam surfaces mounted to the stationary engine structure interact with the toggle mechanism on the connecting rod to shift the toggle between the two stable positions and thereby initiate a change in rod length and compression ratio in response to an engine control signal. It is obvious, however, that other means, e.g. hydrostatic fluid interaction employing a hydraulic plunger, hydrodynamic fluid interaction employing oil nozzles, electromagnetic interactions and hybrid approaches may be used to shift the toggle in response to an engine control signal without departing from the spirit of the invention. 
     
    
     
       DESCRIPTION OF DRAWINGS 
         [0039]    The appended claims set forth those novel features that characterize the invention. However, the invention itself, as well as further objects and advantages thereof, will best be understood by reference to the following detailed description of preferred embodiments. The accompanying drawings, where like reference characters identify like elements throughout the various figures. “Front” views are in the direction of the crankshaft axis wherein the crankshaft is rotating in the clockwise direction, and “rear” views are in the direction of the crankshaft axis wherein the crankshaft is rotating in the counterclockwise direction. The term “oil” indicates any liquid lubricant. 
           [0040]      FIG. 1  is a front view of a piston, cylinder, connecting rod and crank assembly incorporating the inventive two-stage variable compression connecting rod mechanism in the high compression setting: 
           [0041]      FIG. 2  is a rear view of a piston, cylinder, connecting rod and crank assembly incorporating the inventive two-stage variable compression connecting rod mechanism in the high compression setting: 
           [0042]      FIG. 3  is a front view of a piston, cylinder, connecting rod and crank assembly incorporating the inventive two-stage variable compression connecting rod mechanism in the low compression setting: 
           [0043]      FIG. 4  is a front view of a piston, cylinder, connecting rod and crank assembly incorporating the inventive two-stage variable compression connecting rod mechanism in the low compression setting: 
           [0044]      FIG. 5  contains two front sectional views of a piston, cylinder, connecting rod and crank assembly incorporating the inventive two-stage variable compression connecting rod mechanism in the low and high compression settings: 
           [0045]      FIG. 6  contains multiple front views illustrating the transition of the inventive two-stage variable compression connecting rod mechanism from the high compression setting to the low compression setting: 
           [0046]      FIG. 7  contains multiple front views illustrating the transition of the inventive two-stage variable compression connecting rod mechanism from the low compression setting to the high compression setting: 
           [0047]      FIG. 8  provides cutaway perspective views of the eccentric, clutch and toggle mechanism of the inventive mechanism in the low and high compression settings: 
           [0048]      FIG. 9  is an exploded perspective view of the eccentric, clutch and toggle of the inventive mechanism, particularly illustrating a preferred embodiment comprising an array of multiple axial grooves formed in the cylindrical journal bore in the connecting rod small end: 
           [0049]      FIG. 10  contains two phantom front views providing further detail on the component positions of the inventive mechanism in the high and low compression settings: 
           [0050]      FIG. 11  is an exploded perspective view of the connecting rod associated components of the inventive mechanism: and 
           [0051]      FIG. 12  illustrates the relationship between the eccentric and rod geometry, applied loads, frictional loads and resultant eccentric torques. 
       
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
       [0052]    Upon examination of the following detailed description the novel features of the present invention will become apparent to those of ordinary skill in the art or can be learned by practice of the present invention. It should be understood that the detailed description of the invention and the specific examples presented, while indicating certain embodiments of the present invention, are provided for illustration purposes only. Various changes and modifications within the spirit and scope of the invention will become apparent to those of ordinary skill in the art upon examination of the following detailed description of the invention and claims that follow. 
         [0053]    The prior art and the invention are described with reference to four cycle internal combustion engines, but it is to be understood that the invention is applicable to other reciprocating engines, pumps and compressors that might benefit from variable compression. In the description “upper”, “top”, “above” and “head” refer to the direction towards the combustion chamber, and “lower” and “downward” refer to the direction towards the crankcase. 
         [0054]      FIG. 1 ,  FIG. 2 ,  FIG. 3  and  FIG. 4  show a variable compression engine assembly incorporating the two-stage variable length connecting rod of the invention. Piston  100 , cylinder bore  101  and crankshaft  102  incorporating crankpin  103  are of conventional design. The crankshaft  102  rotates within journal  104 , and journal  104  and the cylinder bore  101  are rigidly connected in a fixed geometric relationship by the usual crankcase and engine block structure. They cylinder head (not shown) abuts the upper end  105  of the cylinder bore  101  in a fixed sealing relationship. The two-stage variable length connecting rod assembly  106  connects the crankpin  103  and the piston wrist pin  107 , and like a conventional connecting rod couples the rotary motion of crankshaft  102  to the linear oscillating motion of the piston  100  within the bore  101 . The effective length of the rod  106  is the distance between the centers of the crankpin  103  and the piston pin  107 , and is adjusted by rotating the eccentric  108  between two positions in a bore  109  in the upper end of the rod body  110 .  FIG. 1  and  FIG. 2  show the rod assembly  106  in the maximum length or high compression configuration, and  FIG. 3  and  FIG. 4  show it in the minimum length or low compression configuration. The eccentric  108  is retained in the high compression configuration by ratchet pawl  200  engaging ratchet teeth  201  formed in the outer diameter of the eccentric as illustrated in  FIG. 2 . Similarly, the eccentric  108  is retained in the low compression configuration by ratchet pawl  300  engaging ratchet teeth  301  formed in the outer diameter of the eccentric as illustrated in  FIG. 3 . Ratchet pawls  200  and  300  are connected by a shaft  112  rotating in a journal in the rod body  110 , and form an assembly that rotates as a unit. 
         [0055]    The toggle comprises a first member  202  adjoining the rear face of rod body  110  and a second member  113  adjoining the front surface of the rod body. Toggle members  202  and  113  are connected by a shaft  114  rotating in a journal in the rod body  110 , and form an assembly that rotates as a unit. A member  115 , e.g. a roller, is carried between toggle members  202  and  113 . The member  115  engages a beam spring  116  attached at one end by means of a mount  117  to the rod body  110 . The beam spring  116  is preloaded to apply a force  118  to the member  115  such that a torque is applied to the assembly of toggle members  202  and  113  rotating about the center of shaft  114 . The effect is to form a bi-stable toggle mechanism that takes a first stable position corresponding to a high compression setting as illustrated in  FIG. 1  and  FIG. 2 , and a second stable position corresponding to a low compression setting as illustrated in  FIG. 3  and  FIG. 4 . In the first stable position the toggle member  202  bears against ratchet pawl  200  such that it is engaged with eccentric ratchet teeth  201 , while ratchet pawl  300  is disengaged from eccentric ratchet teeth  301 . Similarly, in the second stable position the toggle member  113  bears against ratchet pawl  300  such that it is engaged with eccentric ratchet teeth  301 , while ratchet pawl  200  is disengaged from eccentric ratchet teeth  201 . In a preferred embodiment the ratchet pawls  200  and  300  and the mating eccentric ratchet teeth  201  and  301  have engagement angles such that the teeth remain engaged when the rod is highly loaded, even after the toggle is flipped and the toggle force is biased to disengage the teeth. The teeth only disengage when the rod force reverses and the eccentric torque passes through zero. After disengagement in one direction, the applied toggle force biases the pawls  200  and  300  so that the eccentric  108  rotation can take place in the opposite direction. These characteristics allow the toggle to be flipped between the two stable positions and thereby pre-set at any point in the engine cycle, and only complete the action when the conditions are favorable. 
         [0056]    The toggle member  113  further comprises a cam follower  119  that extends in front of the rod body  110  and engages control cam  120  during a portion of the engine cycle to flip the bi-stable toggle mechanism from one stable position to the other stable position and trigger a compression ratio change. The control cam  120  is supported by a cam carrier  121  that is free to move a distance  122  in response to a compression ratio change command from the engine control module. 
         [0057]    The cutaway views of  FIG. 5  illustrates further details of the inventive connecting rod assembly  106  in the high and low compression settings at top dead center. Stop pin  500  disposed in stop groove  501  limits the rotation of eccentric  108 . In the high compression setting stop pin  500  carries the eccentric reaction torque resulting from tensile connecting rod loads, and in the low compression setting it carries the torque resulting from compressive rod loads. In each setting the pawls (not visible in  FIG. 5 ) carry the torques in the opposing directions. Toggle shaft  114  and pawl shaft  112  pass through journals in rod body  110 , and a pin  502  secures the beam spring  116  in the spring mount  117 . 
         [0058]      FIG. 6  shows the sequence of events during a transition from the high to the low compression setting. 
         [0059]    In  FIG. 6A  the control cam  120  and the control cam carrier  121  is moved a distance  122  from the high compression position to the low compression position to engage cam follower  119 . This motion is initiated by the engine control module and is carried out by any of a number of known actuation means. The motion may take place any time in the engine cycle, and the cam follower engagement will take place the next time the piston  100  and the rod assembly  106  pass through bottom dead center. 
         [0060]      FIG. 6B  and  FIG. 6C  show the engagement of cam follower  119  with control cam  120 , causing toggle  113  to flip over center. If the motion  122  is incomplete during the bottom dead center event, the toggle  113  may be partially rotated but return to its initial position rather than flip to the new position. In this case the flip will take place during a subsequent bottom dead center event after the control cam  120  has moved further. Once the toggle  113  has flipped over center, the follower  119  no longer contacts the control cam. 
         [0061]      FIG. 6D  shows the pawls  200  and  300  rotated by the toggle  113  to the new low compression setting prior to any rotation of eccentric  108 . This rotation from the position shown in  FIG. 6C  can only take place when the rod load transitions from tension to compression and the torque load on pawl  200  reaches a low level. 
         [0062]      FIG. 6E  shows partial rotation of eccentric  108  driven by compressive load on the rod assembly  106 . The teeth of pawl  300  are forced into engagement with eccentric teeth  301  by the toggle  113 . This forms a one-way clutch that allows forward eccentric rotation under compressive rod assembly loads, and prevents reverse rotation under tensile loads, thereby capturing the partial rotation. 
         [0063]      FIG. 6F  shows full rotation of the eccentric  108  to the low compression setting the after one or more additional compressive load events. In this steady-state condition eccentric compressive reaction torques are carried by the stop pin  500  shown in  FIG. 5  B, and the tensile reaction torques are carried by pawl  300 . In this condition the rod assembly  106  functions as a conventional rod, and there is no interaction between the assembly and the control cam  120 . 
         [0064]      FIG. 7  shows the sequence of events during a transition from the low to the high compression setting. 
         [0065]    In  FIG. 7A  the control cam  120  and the control cam carrier  121  is moved a distance  122  from the low compression position back to the high compression position to engage cam follower  119 . Again, this motion is initiated by the engine control module and may take place any time in the engine cycle. 
         [0066]      FIG. 7B  and  FIG. 7C  show the engagement of cam follower  119  with control cam  120  during a bottom dead center event, causing toggle  202  (not visible) to flip over center to the high compression setting position. If the motion  122  is incomplete during the bottom dead center event, the toggle  202  may be partially rotated but return to its initial position rather than flip to the new position. In this case the flip will take place during a subsequent bottom dead center event after the control cam  120  has moved further. Once the toggle  202  has flipped over center, the follower  119  no longer contacts the control cam. 
         [0067]      FIG. 7D  shows the pawls  200  and  300  rotated by the toggle  202  to the new high compression setting prior to any rotation of eccentric  108 . This rotation from the position shown in  FIG. 7C  can only take place when the rod load transitions from compression to tension and the torque load on pawl  300  reaches a low level. 
         [0068]      FIG. 7E  shows partial rotation of eccentric  108  driven by tensile load on the rod assembly  106 . The teeth of pawl  200  (not visible) are forced into engagement with eccentric teeth  201  by the toggle  202 . This forms a one-way clutch that allows forward eccentric rotation under tensile rod assembly loads, and prevents reverse rotation under compressive loads, thereby capturing the partial rotation. 
         [0069]      FIG. 7F  shows full rotation of the eccentric  108  to the high compression setting the after one or more additional tensile load events. In this steady-state condition eccentric tensile reaction torques are carried by the stop pin  500  shown in  FIG. 5A , and the compressive reaction torques are carried by pawl  200 . In this condition the rod assembly  106  functions as a conventional rod, and there is no interaction between the assembly and the control cam  120 . 
         [0070]      FIG. 8  provides further details of a preferred embodiment of the invention. A portion of rod body  110  is cut away along faces  801 ,  802  and  803  to reveal otherwise hidden features. 
         [0071]      FIG. 8A  shows the high compression setting wherein pawl  200  is engaged with eccentric teeth  201  and stop pin  500  abuts the end of groove  501 , locking eccentric  108  into the high compression setting. This is the steady state condition shown in  FIG. 6F  in which the cam follower  119  does not make contact with control cam  120 , and the toggle assembly  202  and  113  is loaded by beam spring  116  to hold the pawl  200  in engagement with eccentric teeth  201 . Pawl  200  and pawl  300  are rotationally connected through cylindrical member  803  passing through rod body  110  such that pawl  300  is disengaged from eccentric teeth  301 . 
         [0072]      FIG. 8B  shows the low compression setting wherein pawl  300  is engaged with eccentric teeth  301  and stop pin  500  abuts the opposite end of groove  501 , locking eccentric  108  into the low compression setting. This is the steady state condition shown in  FIG. 7F  in which the cam follower  119  does not make contact with control cam  120 , and the toggle assembly  202  and  113  is loaded by beam spring  116  to hold the pawl  300  in engagement with eccentric teeth  301 . The rotational connection  803  disengages pawl  200  from eccentric teeth  201 . 
         [0073]      FIG. 8  shows pawl  200  and eccentric teeth  201  to be wider than pawl  300  and eccentric teeth  301 . This optional feature reflects the fact that pawl  200  and eccentric teeth  201  carry the eccentric reaction torque resulting from compressive rod loads from high compression and combustion pressures, while pawl  300  and eccentric teeth  301  carry the eccentric reaction torque resulting from the lower tensile rod loads from inertial loads. 
         [0074]      FIG. 9  shows a set of grooves  900  in the lower portion  901  of the bore  109  in the upper end of the rod body  110  that modify the squeeze film interface in the journal bearing between the connecting rod body  110  and the eccentric  108 . This aspect of the invention reduces the peak eccentric reaction torque loads on the eccentric latching mechanism during steady state high compression operation, while facilitating the inertial load driven eccentric rotation from low to high compression. The grooves  900  reduce the shortest oil flow path L from the center of the squeeze film bearing area to the edge according to the analysis presented earlier in this disclosure. Since the squeeze bearing force F s  is proportional to 1/L 4 , reductions in L provide effective means of reducing F s  and minimizing squeeze film low friction transients in the oil film in the journal bearing when the applied load is compressing the oil film towards the lower portions of the bearing having the flow passages. This has the effect of suppressing the squeeze film bearing effect, and assuring that the friction coefficient μ is the higher mixed lubrication friction coefficient and not the lower hydrodynamic coefficient, thus reducing the load on the eccentric latching mechanism during these high compressive load intervals. Transient hydrodynamic squeeze film lubrication is, however, retained during the lower force tensile load intervals, facilitating eccentric rotation with the relatively low peak eccentric reaction torque loads present during the transition from low to high compression. During steady state low compression operation pawl  300  carries these relatively low tensile load torques easily despite undiminished squeeze film lubrication between eccentric  108  and the upper groove-free portion of bore  109 . 
         [0075]      FIGS. 10A and 10  B are phantom views of the inventive rod assembly  106 . In particular they show the relationships between eccentric  108 , rod body  110 , stop pin  500 , pawls  200  and  300 , toggle members  113  and  202 , and beam spring  116  in each steady state locked position. 
         [0076]      FIG. 11  provides further details of the components that comprise the exemplary two stage variable compression connecting rod  106 . The toggle members  113  and  202  are joined by shaft  114  that incorporates square coupling ends  1103  and  1104  that transfer torque to matching holes in the toggle members. Shaft  114  rotates in bore  1108  in rod body  110 , and is secured to toggle members  113  and  202  by upset heads  1105  as part of the assembly process. Pintles  1106  and  1107  at the ends of the roller member  115  rotate in holes  1113  in the toggle members. Cylindrical sections  1114  and  1115  incorporated in pawls  200  and  300  rotate within bore  1109  in rod body  110 , and together with bore  1109  comprise the journal bearings that carry the load applied to the pawls by the eccentric  108 . A square extension  1100  of cylinder  1114  couples with a square hole in pawl  200  to transfer torque between pawls  200  and  300  and cause them to rotate in unison. Upset head  1101  formed as part of the assembly operation joins and secures the pawls. Beam spring  116  is inserted into the slot  1112  of spring mount  117  and secured by pin  502  pressed into hole  1110  in the beam spring and  1111  in the rod body  110 . Eccentric  108  is secured in bore  109  of the rod body  110  by pressing stop pin  500  into hole  1116  such that it extends into stop groove  501 . 
         [0077]      FIG. 1  through  FIG. 11  and the accompanying description are primarily intended to illustrate the conceptual features of the invention, and it will be obvious to those skilled in the art that a number of equivalent mechanism details and construction methods can be used to implement the concept. The mechanism details and construction methods, however, are depicted with configurations and proportions adapted to practical internal combustion engines. The basic piston and connecting rod geometry are based on typical modern components in a high speed spark ignition engine with 75 millimeter bore diameter and 65 millimeter stroke. The components added and component modifications needed to realize the two-stage adjustable connecting rod fit within the available space and have proportions consistent with load carrying and dynamic response requirements: 
         [0078]    The rod length change is 4 millimeters, sufficient to vary the compression between 8:1 and 16:1; 
         [0079]    The beam spring orientation parallel to the rod reduces inertial effects on the spring load transmitted to the toggle mechanism; 
         [0080]    The control cam and carrier fit in available space; 
         [0081]    The beam spring is sized to apply about 100 Newtons to the toggle mechanism, which is sufficient to provide calculated toggle and pawl transitions on the order of a few milliseconds, 
         [0082]    The toggle and pawl assemblies are approximately balanced about their rotational axes to minimize response to inertial forces; and 
         [0083]    The pawls, stop pin and eccentric dimensions are selected to withstand the expected mechanical loads.