RIGID CRANKSHAFT CRADLE AND ACTUATOR

Abstract: Poor full power engine performance in variable compression ratio engines resulting from small valve overlap periods, necessary for preventing piston-to-valve strike at high compression ratio, is prevented by phase shifting of the intake and exhaust valves with change of compression ratio. The crankshaft (306) is rotatably mounted in eccentric main bearing supports (310) for adjusting the position of the crankshaft rotational axis (A) relative to the engine housing (302). A drive gear (14) is mounted on the crankshaft (306) and a first driven gear (18) is mounted on a secondary shaft (330) mounted in the engine housing (302). A second driven gear (344) is mounted on a third shaft (346), the second driven gear (344) being in mesh with a drive gear (14) mounted on the crankshaft (306). Rotating the eccentric main bearing supports (310) and moving the crankshaft rotational axis (A) away from the cylinder head (322) of the engine (300) lowers the compression ratio of the engine.

Description: RIGID CRANKSHAFT CRADLE AND ACTUATOR UNITED STATES APPLICATION REFERENCE This application relates to United States Application Number 09/406.124 having a filing date of 09/27/99 BACKGROUND OF THE INVENTION The present mvention relates to a method and apparatus for adjusting the compression ratio of internal combustion engines, and more specifically to a method and apparatus for adjusting the position of the crankshaft with eccentπc crankshaft main bearing supports Designs for engines havmg eccentric crankshaft mam bearmg supports have been known for some time In these engmes the eccentπc mam bearmgs are rotated to adjust the position of the crankshaft's axis of rotation Poor rotational alignment of the eccentπc mam beaπng supports is a problem for these engines because even small amounts of mam bearmg misalignment can cause rapid main bearmg failure Significant forces bear down on the eccentπc mam beaπng supports during operation of the engine In modem passenger car engmes main bearmg loads can exceed 50 MPa The forces exerted on the eccentric main bearing supports are, at times, significantly diflerent from one eccentric mam bearmg support to the next For example, in multi-cylinder engmes a clockwise torque may be applied on a first eccentric main beaπng support from the combustion pressure bearmg down on the first piston, connecting rod and crank throw, and a counterclockwise torque may be applied on a second or third eccentric mam beaπng support from the mertial forces of the second piston and connecting rod pulling up on the second crank throw As a second example, in a smgle cylinder engine having two eccentric mam beaπng supports the torque applied to the crank throw and the resistive torque at the power take-offend of the crankshaft cause uneven loading on the eccentric mam beaπng supports These large unequal forces are a problem because they cause the eccentπc sections to rotate out of alignment with one another causmg rapid failure of the crankshaft mam bearmgs In US patent 887.633, and m German patent DE 3644721 Al a pinned linkage is show for adjusting the rotational alignment of the eccentπc mam bearmg sections US patent 4,738.230 shows dowels extending from each eccentπc mam beaπng support that are fitted mto slots located m a s dable bar for adjustmg the rotational alignment of the eccentric main bearmg supports US patents 5,572,959 and 5.605.120 show gear teeth extending from eccentπc mam beaπng supports that engage a layshaft with matmg gears for adjustmg the rotational alignment of the eccentπc mam bearmg supports US patent 1 ,160,940 shows a bail shaped frame that connects adjacent eccentric sections for adjustmg the rotational alignment of the eccentric sections Poor alignment of the mam bearmgs is a significant problem for each of these systems In addition to poor mam bearing alignment, a number of these systems are not mechamcallv functional for other reasons, are unpractical for mass production manufacture and assembly, and/or are not functional for engines havmg more than two mam bearmgs For example, US patent 1,160,940 shows a bail shaped frame that is weakly connected to the eccentrics and that does not have a rigid construction In addition to not πgidly hold the bearmgs m alignment, the system is not mechamcally functional because the connecting rod does not clear the bail shaped frame The system is also not functional for engines havmg more than two mam beaπngs because it is not possible to slide the eccentric main beaπng support onto the center crankshaft journal or journals A further problem with engmes havmg rotatable eccentπc main bearmg supports in a fixed engine housing is that the location of the crankshaft rotational axis changes with change of compression ratio, making use of a conventional in-line clutch impossible Geared power take-off couplmgs for engmes havmg an adjustable crankshaft rotational axis are shown m the pπor art, however a problem with these systems is that heavy structural reirtforcmg is required to πgidly hold the gear set in alignment In addition to the problem of added weight, engme housmg length is also increased German patent DE 3644721 Al shows a gear set mounted to the free end of one of the eccentric crankshaft main bearmg supports The gear set has an intermediary shaft and an output shaft The output shaft pomts generally away from the crankshaft, and has a fixed axis of rotation for all compression ration settings A problem with the system shown m German patent DE 3644721 Al is that during periods of high engine torque the end eccentπc mam beaπng support may bend out of alignment, resulting in damage to the crankshaft mam bearmg The gear set is also bulky and mcreases cranktrain friction losses due to the increased number of bearings and gear friction US patent 4,738,230 shows a first spur gear mounted on the crankshaft and a second spur gear havmg an axis of rotation that is concentric with the axis of rotation of the mam bearmg supports These gears are too small to carry the torsional loads of the engme US patent 4,738,230 also shows a power take-off system havmg an internal or annular gear set Heavy and lengthy structural reirtforcmg is required for holding the ring gear shaft in rigid alignment with the gear mounted on the end of the crankshaft US patents 5,443.043, 5,572,959 and 5,605.120 show a crankshaft havmg a fixed axis of rotation and an upper engme that changes position relative to its supporting frame when the compression ratio is changed While a conventional m-lme clutch can be employed with this arrangement, the position of the upper engme is changed when the compression ratio is changed, and the inertial mass of the upper engme prevents rapid adjustment of compression ratio A further problem with variable compression ratio engmes is that the exhaust valves must be closed early and the mtake valves opened late m order to prevent valve to piston strike near top dead center (TDC) of the piston The short valve overlap peπod where both valves are open is a problem, because air flow into the engme is restricted causing a loss of engme power Vahe pockets can be formed m the piston to increase valve to piston clearance, however, the pockets add volume to the combustion chamber causing the compression ratio of the engine to be reduced The base height of the piston can be raised further to compensate for the increase of chamber volume, however increasing the piston height mcreases the depth of the valve pockets A significant problem is that the relatively large valve pockets cause increased heat loss from the combustion chamber due to the increased chamber surface area and due to the jagged chamber surface shape The increased heat loss adversely effects engme fuel economy and power Camshaft phase shifters such as those used on the Lexus LS 400, and/or cam profile switching devices such as those used on Honda VTech engmes can be employed to prevent piston to valve strike, however, m addition to bemg expensive, these devices may fail to react fast enough m some vehicles that have been aged Compression ratio may be changed m less than one second, and possibly within a tenth of a second Failure of the vaπable valve device to respond at least as quickly as the vaπable compression ratio devise could result m valve to piston strike, causmg major engme failure resulting in a significant warranty cost A problem with vaπable compression ratio mechanisms is that the actuator consumes a significant amount of energy, off-setting the fuel economy benefit of the variable compression ratio Patent 5 ,611 ,301 issued to Per Gillbrand and Lars Bergsten of Saab Automobile Akhebolag, for example, shows a vaπable compression ratio mechanism where the entire upper engme moves A significant amount of power would be consumed to rapidly move the engme and change the compression ratio, off-setting the fuel economy benefit of the vaπable compression ratio A central problem with vaπable compression ratio mechanisms is the power consumed in the process of adjustmg the compression ratio Primary engme balancing can be accomplished with twin counter rotatmg balance shafts A problem with balance shafts, however, is that of added beaπng friction and windage, which adversely effects engine efficiency and vehicle fuel economy Single balance shaft are employed in many single-cylinder motorcycles such as the Honda XR650L, the Kawasaki KLX250R, and the BMW F650 In these engmes half of the balance mass on the balance shaft and half of the balance mass is placed on the crank web, however, vibration remams significant due to the moment remaining between the crankshaft and single balance shaft axes of rotation SUMMARY OF THE INVENTION In the present invention, a rotatable rigid crankshaft cradle is employed for holding the crankshaft mam bearings in alignment The crankshaft cradle is rotatably mounted m the engine on a pivot axis, and the crankshaft is mounted m the crankshaft cradle on a second axis off-set from the pivot axis An actuator rotates the crankshaft cradle and adjusts the position of the crankshaft axis of rotation and the compression ratio of the engme The crankshaft cradle rigidly holds the mam bearmgs m precise alignment at all tunes and provides long bearing life The crankshaft cradle provides ngid support of crankshafts for smgle and multi- cylinder engmes, ranging from crankshafts havmg two main beaπngs for smgle and two cylmder engmes, to crankshafts havmg five or more main bcaπngs for in-line-four cylinder engmes, V8 engmes, as well as other engines In addition to providing a long main bearmg life, the variable compression ratio mechanism of the present mvention is reliable and has a low cost Referring now to Figs 3, 4 and 5, in an embodiment of the present invention a crankshaft cradle 60 is rotatably mounted m the engme housing on a pivot axis E, and a crankshaft 61 is mounted in the crankshaft cradle on a second axis A off-set from the pivot axis The cradle includes two or more mam beaπng supports or eccentπc members 62 and structural webbmg 64 for rigidly holding the eccentπc members and mam beaπngs in alignment One or more beaπng caps 68 are fastened to the cradle with bolts or another type of fastener for securing the crankshaft in the cradle The bearing caps are removable from the cradle permitting assembly of the crankshaft m the cradle Operation of the mam beaπngs without failure requires precise alignment of the mam bearmg supports at all tunes Accord ng to the present invention, adjacent mam beaπng supports are held in rigid alignment at all times by structural webbmg 64 More specifically, the structural webbmg holds the mam bearmg supports m rigid alignment at all tunes providmg a long service life for the mam beaπngs FIG 9 shows a second embodiment of the present invention As shown m FIG 9, crankshaft cradle 146 mcludes a first eccentric member, or mam beaπng support 160 and a second eccentric member, or mam beaπng support 162 The crankshaft cradle is assembled by sliding main bearing support 160 over a first end of crankshaft 152, and sliding the second mam bearmg support 162 over the second end of crankshaft 152, and πgidly fastening the main beaπng supports together with one or more bolts 164 The mam beaπng supports mclude structural webbmg for rigid attachment of the first mam beaπng support to the second mam bearing support The crankshaft applies large loads on mam bearmgs 12, and the assembled crankshaft cradle 146 holds mam beaπngs 12 m precise alignment under the high load conditions, and more generally crankshaft cradle 146 holds mam bearmgs 12 m precise alignment at all times An actuator first adjusts the rotational position of the crankshaft cradle about its pivot axis, and then locks the rotational position of the cradle m place The actuator applies force on the cradle at a central location between the mam bearmgs. and more generally between the front and back eccentric members, whereby twisting of the crankshaft cradle and miss alignment of the main bearmgs is minimized Accordingly, the eccentπc members are πgidly mamtamed m alignment providmg a long mam beaπng life Another advantage of the present invention is that the cradle has a small inertial mass, and the actuator can adjust compression ratio settings rapidly According to the present invention actuator power is greatly reduced by employing the downward force of on the crank pin to lever up the crankshaft mam bearmgs Specifically, gas force during the power stroke bears down on the crank pin actmg through the piston and connecting rod The crank pin has an orbital diameter Power take-off from the crankshaft is through a drive ear havmg a pitch diameter The gear mesh has a rotational direction pomtmg generally awav from the piston, and applymg a resistive torque on the drive gear proportional to engme power output The pitch diameter of the drive gear is smaller than the orbital diameter of the crank pin, and at approxunately 90 crank angle degrees after to dead center, the gear mesh is located approximately between the crank pm and the crankshaft axis of rotation The downward force of the power stroke actmg on the crank pm is reacted by an upward force proportional to the torque of the gear mesh The downward force on the crank pm and the upward force of the gear torque produces an upward force of the crankshaft on the crankshaft mam bearmgs mounted in the crankshaft cradle A ratchet is employed for ratcheting of the crankshaft cradle and movement of the crankshaft rotational axis towards the cylmder head m steps The present mvention enables compression ratio to be changed with effectively no or almost no power loss to an actuator Additionally, the present mvention has a low cost and exceptional reliability Power is transferred from the crankshaft to the power take-off shaft through gears 14 and 18 Accordmg to the present mvention, gears 14 and 18 have a vaπable centerlme distance and a variable backlash value According to the present mvention, the power take-off shaft is positioned to provide a small maximum gear backlash value for a large change m compression ratio The power take-off couplmg of the present invention provides long gear life, exceptional reliability, low noise levels, and a low cost According to the present invention, the power take-off shaft is located within + 45° of an imaginary first plane and preferably within + 33° The first plane passes through the crankshaft cradle pivot axis E and is perpendicular to the translation axis or centerlme axis of the pιston(s), providmg a small change in backlash from one compression ratio settmg to the next More specifically, location of the power shaft within + 45° of the first plane, and preferably within + 33°, provides a small gear backlash, low gear noise, and long gear life Additionally, gears 14 and 18 are mounted on parallel shafts and preferably have helical mvolute teeth permitting operation of the gears with small variations in centerlme distance Gears 14 and 18 are of automotive quality and have a diameter and width that provides a long gear life Prolonged operation of gears 14 and 18 without failure requires maintenance of parallel alignment of gear 14 and gear 18 Accordmg to the present mvention, the crankshaft cradle holds the beaπng elements, the crankshaft, and gear 14 m precise parallel alignment at all times with the power take-off shaft and gear 18 Accordmg to the present mvention, high structural loads are applied by the crankshaft on the beaπng elements, and the crankshaft cradle rigidly holds the main bearmg supports m precise parallel alignment at all times preventing failure of the bearmg elements and preventing failure of gears 16 and 18 The power take-off shaft is located adjacent to crankshaft cradle m the engine housing, and is rigidly supported with only a rrunimal mcrease of engine size and weight A further advantage of the present mvention is that the power take-off shaft may also serves as a balance shaft FIG 9 shows an embodiment of the present mvention where the power take-off shaft also serves as a balance shaft for the engine The engme shown m FIG 9 has a small size and low bearmg and gear friction, m part because balancing and power take-off is accomplished with a smgle shaft Referπng now to Figs 3, 4. 5 and 9, gear 14 mounted on the crankshaft transfers power from the crankshaft to a second gear 18 mounted on the po er take-off shaft mounted in the engine housing The crankshaft rotates on axis A and the power take-off shaft rotates on axis P Axis A and axis P are separated by a centerlme distance According to the present mvention. rotation of the crankshaft cradle on the pivot axis E adjusts the position of the crankshaft, adjusts the compression ratio of the engine, and changes the centerline distance between axis A and axis P, causing the backlash clearance between gear 14 and gear 18 to change. According to the present invention, a small maximum gear backlash value is provided by locating the axis of rotation of gear 18 on or near a plane that passes through the axis of rotation of the crankshaft and that is generally perpendicular to the line of translation or centerline of the first piston(s). The power take-off arrangement according to the present invention is significantly smaller, lighter, and less costly than prior art systems for engines having eccentric main bearing supports. Additionally, the present invention provides a low friction, compact, and light weight combined balance shaft and power takeoff gear set. The variable compression ratio mechanism according to the present invention holds the crankshaft main bearings in rigid alignment and provides a long bearing life. More specifically, the rigidity of the crankshaft cradle holds the bearings in alignment and prevents damage caused by bearing misalignment and vibration. The present invention is reliable and durable. The present invention can be manufactured using standard materials and mass-production methods, and has a low cost. Another advantage of the present invention is that the main bearings can be line bored, according to current manufactiiring practices, to establish precise main bearing alignment. The variable compression ratio mechanism has a small inertial mass and a fast response providing rapid change of compression ratio. According to the present invention, lowering compression ratio causes the intake valves to open earlier and the exhaust valves to close later, enabling valve to piston strike to be avoided at high compression ration, and enabling high engine power levels to be achieved at low compression ratio. According to the present invention, a drive gear on the crankshaft is in mesh with a driven gear on a second shaft. The two gears are in mesh and have a mesh direction pointing generally away from the piston. The secondary shaft drives a camshaft drive that opens and closes the intake valves. Lowering the compression ratio rotates the driven gear forward causing the intake valves to open earlier. Similarly, a drive gear on the crankshaft is in mesh with a driven gear on a third shaft. The two gears are in mesh and have a mesh direction pointing generally towards the piston. The crankshaft is located between the second and third shafts. The third shaft drives a camshaft drive that opens and closes the exhaust valves. Lowering the compression ratio rotates the driven gear on the third shaft backwards causing the exhaust valves to close later, causing the valve overlap period between the intake and exhaust valves to be increased, resulting in increased engine power. The present invention prevents valve to piston strike at high compression ratio settings. A further advantage of the present invention is that the response rate of valve phase shifting does not deteriorate with engine aging. The present invention is exceptionally robust and reliable. A further advantage of the present invention is that it is significantly less expensive than currently available variable valve control devices. Yet another advantage of the present invention is that it does not have actuator power losses. According to the present invention, primary engine balancing is accomplished with a single balance shaft by off-setting the cylinder axis towards the primary balance shaft. Shifting the cylinder centerline axis towards the balance shaft reduces the off-set moment arm and significantly improves single balance shaft balancing of primary forces. According to the present invention, during the power stroke the crank pin rotates down between the crankshaft axis of rotation and the balance shaft axis of rotation, providmg reduced frictional losses of the piston on the c lmder bore, resultmg in improved fuel economy and mcreased power The present invention is of mimediate benefit to smgle cylmder motor cycle engmes, with improved balance and power at no added cost after tooling an absolute certainty BRIEF DESCRIPTION OF THE FIGURES Fig 1 shows a section of the vaπable compression ratio engme accordmg to the present invention Fig 1 also shows sectional view Fl-Fl of Fig 2 Fig 2 shows sectional view F2-F2 of Fig 1 Figure 2 shows the crankshaft, cradle, power shaft, and power output coupling Fig 3 sho s a three cylmder engme accordmg to the present invention Fig 4 shows a detailed view of the crankshaft cradle, shown in Fig 3 Fig 5 shows a partial sectional view F5-F5 of Fig 3 Fig 6 is a detailed view of fluid chamber 72 shown in Figs 1 and 2 Fig 7 is similar to Fig 1 but shows a two cylmder engme having oil chamber 108, oil chamber 110, and crankshaft cradle 112 Fig 8 shows a partial sectional view of an engine according to the present mvention Fig 9 shows a partial sectional view of an engme accordmg to the present mvention Fig 9 also shows a partial sectional view of engine 136 taken along cut lmes F9-F9 shown m Fig 8 Fig 10 shows a partial sectional view of an engine accordmg to the present mvention havmg a first actuator havmg a connectmg arm and a second actuator havmg a connectmg aπn, for adjustmg and retaining the position of the crankshaft cradle Fig 11 is similar to Fig 3 and shows another embodiment of the present mvention Fig 12 shows a partial section of an engme accordmg to the present invention havmg a close spacing between the crankshaft and the power shaft Fig 14 shows a sectional view of an engme according to the present mvention havmg an adjustable valve timing Fig 14b shows a free-body diagram of forces actmg on the crankshaft Fig 14c shows a ratchet havmg hydraulic valves for movement of the crankshaft cradle in steps Fig 15 shows a detailed view of the camshafts shown in Fig 14 Fig 15b shows an mtake camshaft havmg a phase adjuster Fig 17 shows a cross sectional view of a portion of the engine shown in Fig 14 Fig 17b shows a variation of a portion of the engme shown m Fig 17 RIGID CRANKSHAFT CRADLE AND ACTUATOR Figs 1 and 2 show a portion of a vaπable compression ratio engme 2 accordmg to the present invention Engme 2 has a piston 4, a connecting rod 6, a crankshaft 8 havmg a crankshaft rotational axis A and havmg one or more crank throws or cranks 10 having a crank throw centerline B, crankshaft main beaπngs 12, a crankshaft power take-off gear or output gear 14, a power shaft 16 having a power shaft rotational axis P preferably parallel to crankshaft axis A, a power mput gear or power shaft gear 18. a cylmder 20 havmg coolmg means such as a water jacket 22, a housmg 24, a cylinder head 26, one or more mtake valves 28, one or more exhaust valves 30, fuel injection or carburetion means 32, and one or more spark plugs 34 Crank 10 has a stroke 2L equal to twice the distance from axis A to axis B Crankshaft 8 is rotatably mounted m a ridged crankshaft cradle 36 having one or more eccentrics such as eccentrics 3 and 40 Accordmg to the present mvention. engme 2 includes an actuator 42 (shown in Fig 6) for adjustmg the rotational position of a crankshaft cradle 36 on a crankshaft cradle axis or pivot axis E. and for adjustmg the position of crankshaft rotational axis A relative to housmg 24 More specifically, the cradle is mounted m the engme for pivotmg relative to the engme about the pivot axis, the pivot axis is preferably substantially parallel to and spaced from the rotational axis of the crankshaft, and the actuator varies the position of the cradle about the pivot axis, and adjusts the compression ratio of the engme According to the present mvention actuator 42 may be a hydraulic actuator, an electro-mechanical actuator, a rotary actuator, a straight hydraulic cylmder actuator, or another type of actuator Preferably, engine 2 is a four-stroke port fuel injected spark-ignition engine Those skilled m the art will appreciate that according to the present invention engme 2 may be a direct fuel injection spark-ignition engme, a diesel engme, a two-stroke engme. or another type of reciprocating piston engme or variable volume machme such as a Stirling engme, a steam engine, a pump, a compressor, or an expander (all not shown), and that other effective arrangements of valving, fuel supply and ignition means may be provided and/or omitted Those skilled in the art will appreciate that housmg 24 and cylmder head 26 may be separable, a smgle cast part, or other functional arrangement Piston 4 is shdably housed m cylinder 20 which is provided air through mtake valve 28 Intake valve 28 may include an adjustable valve actuation mechanism 44 Engme 2 has one or more cylinders 20 In rnulti-cylinder engmes according to the present invention, the cylinders are preferably in-line or m a steep * 'V' orientation, as shown m Fig 7. however other arrangements may be used Referring now to the single cylmder shown in Figs 1 and 2. within engme 2 the geometric cylmder displacement D of the cylinder within engme 2 is equal to the product of the full stroke of piston 4 in cylmder 20 t ies the cross sectional area of cylmder bore 20 The engme displacement or cylmder displacement of engines accordmg to the present mvention havmg one or more c lmders is the sum of the geometric cylmder displacements of all of the working cylmders of the engme Imaginary pomt X is located at the geometπc center of the cross sectional area of cj lmder bore 20. and immediately above (just out of reach of) piston 4 when piston 4 is fully extended away from crankshaft rotational axis A when engine 2 is at its highest compression ratio settmg Preferably, cylmder bore 20 is round, however cylinder bore 20 may have other cross sectional area shapes such as oval, square, or another shape Those skilled m the art will appreciate that the top of piston 4 may be flat or have a non-flat surface The cylmder within engme 2 has a combustion chamber volume, or end chamber volume, d havmg a minnnum d mm and a maximum d,™ * Combustion chamber volume d is the volume between cylmder head 26 and piston 4 when piston 4 is fully extended away from crankshaft rotational axis A Crankshaft rotational axis A has a first position located on an axis F that provides the smallest combustion chamber volume, d, Combustion chamber volume ά,^ is the volume between cylmder head 26 and piston 4 when piston 4 is fully extended away from crankshaft rotational axis A and crankshaft rotational axis A is located on axis F (e g , rotational axis A is at its closest position to imaginary pomt X) Crankshaft rotational axis A has a second position located on an axis G that provides the largest combustion chamber volume, d m>x Combustion chamber volume d m a x is the volume between cylmder head 26 and piston 4 when piston 4 is fully extended away from crankshaft rotational axis A and crankshaft rotational axis A is located on axis G (e g , rotational axis A is at its farthest position from imaginary pomt X) The compression ratio C of the cylmder shown within engme 2 is equal to, C = (D + d)/d The maximum compression ratio C π ^ of the cylmder shown within engme 2 is equal to, The minimum compression ratio C mm of the cylmder shown within engme 2 is equal to, Crankshaft cradle 36 is rotatably mounted m a bore 46 m housmg 24 Crankshaft cradle 36 mav haλ e a first eccentπc member, main beaπng support or section 48 and a second eccentric member, mam bearmg support or section 50 Crankshaft cradle 36 has one or more eccentrics such as eccentrics 38 and 40 Eccentric 38 is formed m section 48, and eccentπc 40 is formed m section 50 Section 48 includes webbmg 52. and section 50 includes webbmg 54 Webbing 52 and 54 πgidh connects eccentπc members 48 and 50 to one another In detail, eccentrics 38 and 40 are rigidly jomed by webbmg 52 and 54, and may be held in position by a fastener such as pm. clip, screw or bolt 56 and more generalh eccentric member sections 48 and 50 are rigidly, and preferably removably, connected together with one or more fasteners Referring now to Figs 3, 4 and 5, in an embodiment of the present invention, crankshaft cradle 60 has eccentric members or sections 62 Adjacent eccentric members or sections 62 are πgidly jomed by webbmg 64 Adjacent eccentric members or sections 62 joined by webbmg may be an single cast part (as shown), or may be an assembly of parts, and more specifically crankshaft cradles compnsrng two or more eccentric members 62 and webbmg 64 may be a one-piece cast part or an assembly of parts Figs 4 and 5 show four eccentπc members 62 and webbmg 64 cast together as one πgid part and supporting four mam bearmgs 66 Sections 68 may serve as crankshaft mam bearmg caps Beaπng cap bolts or fasteners 70 rigidly and preferably removably secure said beaπng caps 68 to eccentπc member or sections 62 Referring now to Figs 1 through 5, according to the present mvention, aφacent eccentric members are rigidly jomed by webbmg effective for πgidly holding the eccentπc members and the crankshaft mam beaπngs m alignment on crankshaft centerlme axis A Referring to Figs. 1 and 2, mam beaπngs 12 are mounted or formed m eccentrics 38 and 40 for supporting crankshaft 8 Bearmgs 12 may be journal bearmgs, roller, needle, tapered, spheπcal, or ball beaπngs, or any other functional beaπng means for supporting crankshaft 8 m eccentric 38 and 40 Preferably beaπngs 12 are separable pernuttmg assembly of crankshaft 8 m crankshaft cradle 36 Bearmgs 12 may be separable by sliding sections 48 and 50 apart along axis E Referring now to Figs 3, 4 and 5, bearmgs 66 are separable by removmg bolts 70 and separating eccentπc member or section 62 and bearmg cap or section 68 Referπng now to Figs 1 and 2, crankshaft cradle 36 and eccentπcs 38 and 40 rotate about a pivot axis E According to the present mvention, one or more fluid chambers 72 are formed between housmg 24 (andor housmg 24 plus one or more end surfaces 74 and 76) and crankshaft cradle 36 Those skilled m the art will appreciate that other surfaces may be used to contain fluid within chamber 72 The fluid m chamber 72 is oil or a similar hydraulic working fluid The rotational position of crankshaft cradle 36 and eccentπcs 38 and 40 on pivot axis E is adjusted by adjustmg the volume of chamber 72 Preferably, the fluid in chamber 72 exerts a force directly on crankshaft cradle 36, causmg crankshaft cradle 36 to rotate about pivot axis E, and causmg the position of crankshaft rotational axis A to be adjusted The volume of chamber 72 is adjusted by admitting or releasmg fluid from chamber 72, and m more detail by pumping fluid into chamber 72 or releasmg fluid from chamber 72 Chamber 72 is m fluid communication with one or more fluid passageways 78 One or more valves 80 control flow of fluid through passageway 78 (or other passageway m fluid cornmu cation with chamber 72), and thus control flow of fluid to and out of chamber 72 Valve 80 is controlled by a controller 82 or other control means Crankshaft cradle 36 and eccentrics 38 and 40 may rotate up to θ degrees from a first position to a second position In the first position, crankshaft rotational axis A is located on axis F, and in the second position crankshaft rotational axis A is located on axis G Referring now to the combustion chamber volume d shown within engme 2, releasing fluid from chamber 72 through valve 80 causes crankshaft cradle 36 to rotate (clockwise) about pivot axis E θ degrees (due to downward force on eccentπc sections 38 and 40. and on crankshaft cradle 36 from crankshaft 8 and/or due to other applied forces), causmg crankshaft 8 to move (be lowered) from centerlme F to centerline G, causmg volume d to be mcreased from dm m to d max and causmg the compression ratio C of the cylmder shown withm engme 2 to be reduced from C ma to C m u, Fluid can be pumped back mto the chamber 72 to rotate crankshaft cradle 36 counterclockwise, causmg the compression ratio C to be mcreased Those skilled in the art will appreciate that the crankshaft rotational axis A can be adjusted to any position between axis F and axis G, and compression ratio C can be adjusted to any value between C m∞ and C nm , According to the present mvention, the volume of chamber 72 is adjusted to adjust the rotational position of crankshaft cradle 36 and eccentπcs 38 and 40 Adjustmg the rotational position of crankshaft cradle 36 and eccentπcs 38 and 40 adjusts the position of crankshaft rotational axis A (e g , the rotational centerlme position of crankshaft 8) relative to housmg 24, and adjusts the compression ratio C of engine 2 Those skilled m the art will appreciate that engme 2 can have one or more cylmders, and that the compression ratio C, displacement D, and combustion chamber volume d can be the same or different for each of the cylinders accordmg to the present mvention Fig 6 shows a detailed view of chamber 72. and more generally a rotary actuator 42 for rotatmg crankshaft cradle 36 and eccentrics 38 and 40 relative to housmg 24 Refemng now to Figs 1, 2 and 6, crankshaft cradle 36 has a surface 84 at radius Rl from pivot axis E that slidably engages a first chamber end surface 86 extending from bore 46 Surface 84 is preferably located on webbmg 52 and 54 Those skilled in the art will appreciate that surface 84 may touch end surface 86. or be separated from end surface 86 by a small clearance (e g , by a small working tolerance between parts) Chamber 72 has a second chamber end surface 88 extending from surface 84 that slidably engages bore surface 46 Those skilled in the art will appreciate that end surface 88 may touch bore 46, or be separated from bore 46 by a small clearance (e g , by a small working tolerance between parts) Chamber 72 is formed by surface 84, bore surface 46. end surface 88, end surface 86. and a top surface 74 and a bottom surface 76 Those skilled in the art will appreciate that top surface 74 and/or bottom surface 76 may be a continuation of, or radiused from, surface 46, surface 84, surface 88. and/or surface 86 One or more seals may be used to retain flmd in chamber 72, such as face seals 94 and 96, lme seals 98 and 100, and end surface seals 102 and 104 Those skilled m the art will appreciate that other seal types and arrangements may be used to retam fluid in chamber 72 Accordmg to the present invention, hydraulic fluid in chamber 72 acts on crankshaft cradle 36 More generally, crankshaft 8 is mounted in eccentrics 38 and 40 in crankshaft cradle 36, and crankshaft cradle 36 is the rotary element of rotary actuator 42, e g , crankshaft 8 is mounted in the rotary element of the rotary actuator The present mvention is compact m design and provides ridged support of crankshaft 8, which improves crankshaft durability and life, and reduces vibration and noise The present invention is simple m design and inexpensive to manufacture, and has exceptional reliability and durability At times duπng operation of the present invention, the flmd in chamber 72 is at high pressure, such as duπng the power stroke of engine 2 when piston 4 is beaπng do n on connectmg rod 6 Duπng the intake stroke of engine 2, the downward motion of piston 4 and connectmg rod 6 may cause crankshaft 8 to exert an upwards force on eccentrics 38 and 40. causmg crankshaft cradle 36 to rotate counterclockwise, and the flmd pressure in chamber 72 to decrease Crankshaft cradle 36 may be held m position by retaining means such as a pre-tensiomng spring 106 (see Fig 9), a second hydraulic flmd chamber (see Figs 7, 10, and 12), a friction brake, a sliding pin, or other means that fixes or substantially retain and/or hold firm the position of crankshaft cradle 36 relative to housmg 24 Pre-tensiomng spπng 106 may be used to exert a clockwise torque on crankshaft cradle 36 (e g., sprmg 106 moves crankshaft axis A in a direction generally away from piston 4, encouraging the compression ratio to be reduced), to minimize and/or prevent counterclockwise movement of crankshaft cradle 36 when a change of compression ratio is not being sought Spring 106 minimizes and/or substantially prevents rotational vibration or bounce of crankshaft cradle 36 m bore 46 Fig 7 is similar to Fig 1 except that Fig 7 shows a first flmd chamber 108, a second fluid chamber 110, a crankshaft cradle 112, and webbing 111 Chamber 108 is similar to chamber 72 (shown m Figs 1 and 6) m that increasing the volume m chamber 108 (e g , by pumping hydraulic fluid mto chamber 108) rotates crankshaft cradle 112 counterclockwise, causmg crankshaft 8 to be raised and the compression ratio C to be mcreased Chamber 110 is filled with fluid to retain crankshaft cradle 112 m a fixed or near fixed position, and prevent crankshaft cradle 112 from substantively rotatmg or vibrating under the cyclic (and m some cases reversmg) loads applied to crankshaft cradle 112 by crankshaft 8, and m more detail to retam crankshaft cradle 112 in a fixed or near fixed position except during periods when valves 114 and/or 116 are adjusted to adjust the position of crankshaft cradle 112 and mam beaπngs 12 relative to housmg 24 Chamber 110 may also be used to forcibly rotate crankshaft cradle 112 clockwise, causmg crankshaft 8 to be lowered, and causmg the compression ratio C to be lowered Controller 82 and valves 114 and 116 are used to control feed of flmd into and out of chambers 108 and 110 through fluid passageways 118 and 120 Other valves and flmd passageways, and other valve and flmd passageway aπangements may be used to control the volume of fluid m chambers 108 and 110 Referπng now to Figs 1 and 2, power is transferred from crankshaft 8 to power take-off shaft 16 through a power output coupling 58 compπsmg gears 14 and 18 Accordmg to the present mvention, the distance between the crankshaft rotational axis A and the power shaft rotational axis P changes as the crankshaft rotational axis A is moved and the compression ratio of the engine is changed More specifically, the power output couplmg has at least one external power take-off gear 14 on crankshaft 8 and power shaft 16 has an axis of rotation P and an external power mput gear 18 External power take-off gear 14 is engaged with external power mput gear 18. and crankshaft 8 has a first axial position having a first distance from power shaft axis P at a said first pivot position of cradle 36, and crankshaft 8 has a second axial position havmg a second distance from power shaft axis P at a second pivot position of cradle 36, and the second distance is greater than said first distance Gears 14 and 18 are external gears (not internal or annular gears) and have mvolute, epicycloid or other suitable gear tooth shapes so that the durability of the gears is not substantively effected by minor changes m the centerlme distance between the crankshaft 8 and the power shaft 16 Preferably gears 14 and 18 are helical gears havmg parallel axes of rotation, to provide a higher load carrying capacity, a higher operational speed capability, and reduced noise Referπng now to Figs 1 and 7, each piston 4 in the engme has a translation axis 91 Engines accordmg to the present mvention have a mean translation axis or centerline axis 92, where the centerlme axis 92 is defined as the translation axis 91 in smgle cylinder engmes, and the bisectmg or average translation axis in multi-cylinder V or W engmes In order to minimize change m the distance between the crankshaft gear 14 and the power shaft gear 18 during changes of compression ratio, in present embodiment, axis P is positioned within plus or minus 45° of a first plane Specifically, a first plane 90 passes through pivot axis E and is perpendicular to the centerlme axis 92 A first crankshaft axis is located approximately on the first plane, said centerline axis and said crankshaft axis bemg on the same side of said pivot axis A second plane 90b passes through the first crankshaft axis, said second plane and said first plane bemg separated by 45°, and a third plane 90c passmg through said first crankshaft axis, said third plane and said first plane bemg separated by 45° and said second plane and said third plane being separated by 90° Axis P is located between the second plane and the third plane, thereby mimizmg the maximum backlash between the external power take-off gear and the external power input gear Those skilled m the art will appreciate that axis P may be located to the right or left of crankshaft 8 accordmg to the present invention Alternatively, the first plane has its ongin at, and is perpendicular to. a second plane that passes through axis F and G Axis P is positioned within plus or rninus 45° of the first plane, where the plus or minus 45° is measured from the oπgm of the first plane Those skilled in the art will also appreciate that placement of axis P within plus or minus 45° of the first plane provides a minimum gear backlash m engines both having rigidly connected and not rigidly connected mam bearmg supports An anti-backlash gear 112 may be used to prevent gear chatter and wear Anti-backlash gear 112 is sprmg loaded to keep the larger load bearmg gear 18 in contact with its matmg crankshaft gear 14 at all or almost all times Alternatively, an anti-backlash gear may be mounted on crankshaft 8 Po er shaft 16 may have one or more balance weights 124 Those skilled m the art will appreciate that the balance weight 124 is optional Accordmg the current embodiment of the present mvention, the power output of the engme is through the power shaft, since its centerlme is fixed along axis P, and thus power shaft 16 can easily be coupled to a clutch, transmission or other rotatmg element (all not shown) Power output for boats, airplanes, and some other applications may be directly through crankshaft 8, as adjustmg the centerlme of crankshaft 8 may not significantly affect system performance Referπng now to Figs 1, 2. and 6, preferably the engine is assembled by sliding crankshaft cradle 36 mto bore 46 along axis E Bore 46 m housmg 24 can be machmed at low cost, and provides ridged support of crankshaft cradle 36 One or more parts 126 may be attached (or foπned into the inside of bore 46) b\ a screw 128 or other attachment means such as a bolt, a slot, or adhesive Those skilled in the art will appreciate that other parts may be attached or foπncd into the inside of bore 46 Attaching parts inside bore 46 (as opposed to machining forms extending inward from radius R2) enables bore 46 to be machined at low cost An opening 130 (dashed lines) may be provided for access to bolts and for oil drainage A significant advantage of the present mvention is that crankshaft cradle 36 and housmg 24 πgidly hold crankshaft mam beaπngs 12 m alignment (for smgle and multi-cylmder engmes) Rigidly supporting the crankshaft mam bearmgs 12 m alignment significantly improves crankshaft durability, and reduces noise and vibration Those skilled m the art will appreciate that a crankshaft for a multi-cylmder/piston engme can be rigidly supported with the present mvention, and for example with an eccentric that has more than two πdged crankshaft beaπng supports In the single cylmder engme shown in Figs 1 and 2, crankshaft cradle sections 48 and 50 slide onto the ends of the crankshaft 8, and may also slide mto bore 46 The crankshaft cradle sections 48 and 50 may be fastened together by a screw 56 or by other fastener means such as a bolt, pin, brazmg, or adhesive Preferably, end plates 132 and 134 are bolted to housmg 24 to secure crankshaft cradle sections 48 and 50 in place Endplates 132 and 134 may be used to retain crankshaft cradle sections 48 and 50 m position Boltmg endplates 132 and 134 to housmg 24 may compressively set seals 102 and 104 in place Those skilled m the art will appreciate that one or both endplates may be foπned in housing 24 (for example, one or both end surfaces 76 and 74, may be machmed out of housmg 24), and/or other means may be used to retam crankshaft cradle sections 48 and 50 m position Fig 11 shows m sectional view part of a three cylinder variable compression ratio engine accordmg to the present mvention, havmg a piston 4, a connectmg rod 6, a crankshaft 61 having a rotational axis A and crankshaft beaπngs 66, a cylmder 20. m a housing 59, an crankshaft cradle 60, and an eccentπc mam cap 71 Crankshaft cradle 60 comprises eccentric members or section 62 and webbmg 64 rigidly connectmg two or more of the eccentric members 62 Eccentric members 62 and bearmg caps or sections 68 have a separation surface 63 Those skilled m the art will appreciate that separation surface 63 may be on an imaginary flat plane that bisects axis A, a curved surface that bisects axis A, or another imaginary surface that allows assembly of crankshaft 61 mto crankshaft cradle 60 Sections 62 and 68 are joined by bolts or fastener 70 or other functional means Crankshaft cradle 60 is rotatably supported m housmg 59 by eccentπc mam cap 71 Removable mam cap 71 enables crankshaft cradle 60 to be laid mto the housmg as an alternative to the shde- m assembly described above Specifically, Fig 1 shows a rigid engine construction having a housing 24 havmg an upper housmg portion 24a and a lower structure 24b, where the upper housing portion 24a and the lower structure 24b is a one-piece metal casting, and eccentric members 48 and 50 slide mto housmg 24 on ax s E Bore 46 may be foπned m lower structure 24b, or lower structure 24b may support a bearmg element havmg a bore 46 for supporting the cradle (not shown) Referring now to Fig 11 , an oil feed line 65 m section 62 and an oil supply galley 67 provide oil to crankshaft bearmgs 66 Galley 67 is preferably about as wide as it is deep Referring now to Figs 3 and 5. oil feed line 77 is m webbmg 64 and oil feed line 65 is in eccentric member 62 Fig 4 shows a detailed view of cradle or crankshaft cradle 60 Crankshaft cradle 60 has eccentπc members or sections 62 for πgidh supporting crankshaft bearings 66 Eccentric members or sections 62 are rigidly joined or connected to one another by cross webbing structure 64 Refemng now to Figs 1 and 2, eccentπc members or sections 48 and 50 are rigidly jomed or connected to one another by cross webbmg 52 and 54 Accordmg to the present invention, crankshaft cradle 36 mcludes cross webbing structure 52 and 54 effective for πgidly holdmg crankshaft mam beaπngs 12 in alignment, and crankshaft cradle 60 includes cross webbmg structure 64 effective for rigidly holdmg crankshaft mam bearmgs 66 m alignment Referπng now to Figs 3, 4 and 11, cross webbmg structure 64 has an outer surface 69a that bears on a bore surface in housing 59 mcluάing an inner housmg surface 69b and on an inner main cap surface 69c Crankshaft cradle 60 havmg outer surface 69a is rotatably mounted inside said bore surface m housmg 59 and/or eccentπc mam cap 71 Outer surface 69a may extend onto the outer surface of webbmg structure 64, and outer surface 69a may form a continuous surface between adjacent eccentric members or sections 62 (shown) According to the present mvention, crankshaft cradle 60 may be supported along all or a portion of bearmg surface 69a Fig 8 shows a partial sectional view of an engme 136 according to the present mvention Fig 8 is similar to Fig 1 except that Fig 8 shows a piston type hydraulic actuator 138 havmg a hydraulic piston 140 slidably housed m a hydraulic cylmder 142 for linear translation movement Piston 140 is pivotaly connected to an actuator link or arm 144, and arm 144 is pivotaly connected to a crankshaft cradle 146 Piston 140 may be connected to cradle 146 by actuator link or arm 144 or by another type of coupling such as a rack and pinion gear set, an eccentric bushmg between arm 144 and bolt or pm 164, or another functional arrangement Flmd enters and exits cylinder 142 through one or more passageways 148, and flow of flmd mto and out of cylmder 142 is controlled by one or more valves (not shown) Accordmg to the present mvention, pressurized flmd entering cylmder 142 through passageway 148 forces piston 1 0 and arm 144 in a generally downward direction (with respect to the oπentation of engme 136 shown m Fig 8) causmg crankshaft cradle 146 to rotate counterclockwise about axis E causmg crankshaft centerline A to πse and the compression ratio of engme 136 to be mcreased An actuator first adjusts the rotational position of the crankshaft cradle about its pivot axis E, and then locks the rotational position of the cradle m place Referπng now to Figs 2, 5 and 9, according to the present mvention, the actuator is preferably connected to the middle of the crankshaft cradle, e g , between the front and the back main bearmgs (e , between the two mam bearings that are spaced farthest apart) and more generally between the front and back eccentric members or mam beaπng supports (e g , between the two eccentric members that are spaced farthest apart), providmg a centrally applied force on the cradle, whereby twisting of the crankshaft cradle and misalignment of the mam beaπngs is minimized Fig 9 shows placement of actuator arm 144 between the mam bearmgs 12 between eccentric members 160 and 162, providing balanced loading of actuator force on crankshaft cradle 146 Fig 5 shows placement of actuator arm 144 between the mam bearmgs 66 and more generally between the t o eccentric members 62 spaced farthest apart, providmg balanced loading of actuator force on crankshaft cradle 61 Fig 2 shows the flmd chamber of an actuator 42 applymg even pressure on crankshaft cradle 36 along its length, and more generally between eccentπc members 48 and 50 Accordingly, the eccentπc members are πgidly mamtamed m alignment providing a long mam beaπng life Fig 9 shows a partial sectional view of engine 136 taken along cut lmes F9-F9 shown m Fig 8 Referπng now to Figs 8 and 9, engme 136 has a housing 150, a piston 4, a connectmg rod 6, a crankshaft 152 mounted in bearmgs 12 havmg an inner diameter 154 for carrying crankshaft 152, and beaπngs 12 are housed m crankshaft cradle 146 Hydraulic cylmder 142 is formed m or πgidly aligned with housmg 150 Connectmg rod 6 has a big-end beaπng 156, and is rotatably mounted on crankshaft 152 on crank 158 havmg a beaπng axis B Preferably crankshaft cradle 146 has a first eccentπc section 160 and a second eccentric section 162 that slide onto opposite ends of crankshaft 152, and are πgidly held together by one or more fasteners such as bolts 164 and 166, or by other means such as a pm or screw Section 160 mcludes a first structure 168 for retaining bolts 164 and 166, and section 162 mcludes a second structure 170 for retaining bolts 164 and 166 Bolt 164 may serve as a connectmg pm, linking or pivotaly connecting aπn 144 and crankshaft cradle 146 Preferably, bolt 164 serves as a connecting pin and is generally centered between section 160 and section 162. so that force from arm 144 is substantially applied equally to sections 160 and 162 m order to lnimnuze misalignment of beaπngs 12 In detail, the connectmg pm portion of bolt 164 is located in the axial direction along axis E between sections 160 and 162, and is located in the radial direction outside the swept path of crankshaft 152, connecting rod 6 (including the connecting rod big end beaπng cap), and counterweights (172 shown in Fig 9) Crankshaft 152 may have counterweights 172 In Fig 8. counterweights 172 are not shown (I e cut away) to show beaπng 156 at the big end of rod 6, and the crankshaft mam bearmgs 12 As shown in Fig 9, a pre-tensiomng means m the form of a sprmg 106 applies a torque on crankshaft cradle 146 Spring 106 may be attached directly to crankshaft cradle 146 and housmg 150 Preferably, sprmg 106 is coiled around axis E and attached to an end of crankshaft cradle 146 Referπng now to Figs 8 and 9, spπng 106 exerts a clockwise torque on crankshaft cradle 146. and encourages or causes the compression ratio of engme 136 to be decreased, and more specifically spring 106 exerts a torque on crankshaft cradle 146 causmg (or encouraging) crankshaft cradle 146 to rotate causing crankshaft 152 to move m a direction away from piston 4 (e. , causing or encouraging the compression ratio of engine 136 to be reduced) Hydraulic pressure m cylmder 142 acts against (e g , resists) the torque on crankshaft cradle 146 from sprmg 106. and encourages or causes the compression ratio of engme 136 to be mcreased Oil is fed to beaπngs 12 and 156 through an oil supply fitting 176 preferably located on axis E and havmg an oil feed passageway 178. that is m flmd communication with oil feed l es (e g . crankshaft passageways) 180 and 182 Preferabh oil feed line 180 is located or centered on axis A, supply fittmg 176 is located or centered on axis E, and supply fittmg 176 is attached directly to section 160 An off-set passageway or eccentric transition space 184 connects feed line 180 and oil feed passageway 178 in fittmg 176 Supply fittmg 176 may include a rotary fitting or joint so that oil feed passageway 178 may remam stationary when section 160 and crankshaft cradle 146 rotate During operation of the present invention, oil enters passageway 178 and flows mto off-set passageway 184 The oil then flows to beaπngs 12 and 156 through passageways 180. branch passageway 186, and 182 Those skilled m the art will appreciate that other flmd passageway arrangements may be used according to the present mvention to deliver oil to beaπngs 12 and 156 Surfaces 188 and 190 may be lubricated by feed lme 192 and/or 194 Gear 14 may have a helical or bevel tooth pattern 196 that pushes crankshaft cradle 146 m the direction of fittmg 176 Crankshaft cradle 146 may have or bear on a thrust bearmg 198 that resists axial thrust exerted by gear 14 or other axial thrust forces from other sources Those skilled m the art will appreciate that other types of thrust bearmgs may be used according to the present mvention Gear teeth 196 bearmg down on power shaft gear 18 result m a reactionary upward force on gear 14 and crankshaft 152 The present mvention mcludes a ridged crankshaft cradle 146 and a stiff housing 150 to prevent crankshaft cradle 146 from twisting under these and other forces and loads The crankshaft cradle may be fabricated m cast iron, steel, aluminuin, magnesium, titamum. or another mateπal or combmation of materials to provide πdged support of the crankshaft mam bearmgs Axis B and axis A are separated by length L The stroke of the crank throw is 2L The stroke of engme 136 is approximately 2L, and varies slightly because the cylinder axis does not mtersect the crankshaft axis for all compression ratio settings In general, the stroke of engme 136 is assumed to be 2L. with minor variances m stroke length ignored Referπng now to Figs 7 and 12, duπng the operation of the engine, the movement of the connecting rod defines a connectmg rod swept path The webbmg and eccentric members are located entirely outside the connecting rod swept path to prevent mechanical interference According to the present mvention, the engine may have a clearance zone to minimize crankshaft cradle mass and/or to provide clearance around a balance shaft 200 and/or the power shaft 5 In detail, accordmg to the present invention, each piston has a translation axis, and the engine has a mean translation axis or centerline axis 92, where the centerlme axis is defined as the translation axis m smgle cylmder engmes, and the bisectmg or average translation axis in multi-cylinder V or W engines Engme 258 has a cradle 260. webbing 262, a first plane 90 originating at pivot axis E and passing through centerlme axis 92. the first plane and the centerlme axis bemg peφendicular The engine has a second plane 250 originating at pivot axis E. the second plane bemg separated from the first plane 90 by 20 degrees, and the engine has a third plane 252 originating at pivot axis E, the third plane and the first plane being separated by 20 degrees, and said second plane 250 and said third plane 252 being separated by 40 degrees The connectmg rod swept path is bound by a fourth plane 254 and a fifth plane 256 (see Fig 9), said fourth and said fifth planes bemg peφendicular to the rotational axis of the crankshaft Engme 258 has a clearance zone bound by the second plane 250 and the third plane 252 and bv the fourth plane 254 and the fifth plane 256, and the webbmg 262 is located exclusively outside of said clearance zone at all compression ratio settings, providing a mechanical clearance between the crankshaft cradle and balance shaft 200. power shaft 5. and/or other engme components Referπng now to Fig 9, to provide ridged support of crankshaft beaπngs 12, crankshaft cradle 146 has a maximum thickness t between a first circle or cylmder 147 and a second circle or cylmder 149 The first circle 147 has a center on the rotational axis of the crankshaft A and has a diameter of 1 2 tunes the stroke of the crank throw, and the second circle 149 has a center on the rotational axis of the crankshaft A and has a diameter of 2 0 times the stroke of the crank throw Preferably, the maximum thickness between the first and second circle is at least 0 10 times the thickness of the stroke of the crank throw providing a rigid cradle Preferably, the maximum thickness along the first circle is also at least 0 10 tunes the length of the stroke of the crank throw Accordmg to the present mvention, the ratio of the thickest section t of crankshaft cradle 146, between circles 147 and 149, divided by length L is greater than 0 10, (e g , t/L > 0 10) providmg ridged support of mam beaπngs 12 Similarly, the crankshaft cradle has a second maximum thickness t2 on a plane 151 peφendicular to the rotational axis A of the crankshaft and passmg through the crank throw 158 The second maximum thickness t2 is at least 0 10 times the length of said stroke providmg a πgid cradle (e , t2/L > 0 10) As stated before, the crankshaft cradle may be a one-piece cast part, or an assembly of parts Preferably, the webbing has a first portion, and the first portion has a thickness at a radial distance from the rotational axis of the crankshaft greater than the stroke, wherem a first eccentric member and the first portion is a one-piece metal casting, providing a rigid structure between the eccentric member and the webbmg used tojom adjacent eccentric members To provide low mechanical forces on the cradle, and a high vibrational natural frequency of the cradle (higher than the maximum operational speed of the engine), preferably the distance between the pivot axis and the crankshaft axis is at a immmum Specifically, preferably the pivot axis passes through the swept path of the connecting rod In any event, crankshaft cradle 146 provides ridged support of beanngs 12, and more specifically crankshaft cradle 146 holds beaπngs 12 m alignment within a tight tolerance, where the tight tolerance is small enough to prevent failure of bearmgs 12 or failure of crankshaft 152 In engines according to the present mvention havmg two or more mam bearmg supports, and preferably engmes with journal beaπngs according to the present mvention, the tight tolerance is preferably a radial deflection of less than 0 008 mches (and preferably less than 0 004 mches) of the centerline of any one bearmg 12 from the centerline of crankshaft cradle 146, and more specifically, measured from a zero deflection baseline where crankshaft bearmgs 12 are on a first straight axis of rotation and the crankshaft is on a second straight axis of rotation that is concentric with the first axis of rotation Those skilled in the art will appreciate that the present invention provides a tight tolerance for crankshaft cradles that support crankshafts for one or more cylmder engmes In vehicles (such as m light duty passenger cars and light trucks as defined by the U S Environmental Protection Agenc ) applications of the present mvention. crankshaft cradle 146 has a πgidi great enough to prevent failure of bearings 12 within a minunum of 100,000 miles of vehicle use Light duty passenger car and truck engines are operated at part load most of the time According to the present mvention, beaπng alignment is measured at a first engme settmg havmg a crankshaft rotational speed between 1200 rotations per minute (φm) and 6000 rpm, and at an engme mean effective pressure (mep) of less than 500 kilopascals ( 500 kPa) Mean effective pressure is defined on page 50 of Internal Combustion Engme Fundamentals, John, B Heywood McGraw-Hill Book Company. 1988, as follows, mep(kPa) = P(kW)n R x 10 3 / V d (dm 3 )N(rev/s) n R is equal to two (2) for four-stroke engmes and one (1) for two-stroke engmes V d is swept engme displacement N is engme rotational speed m revolutions per second, and P is power m kilowatts More specifically, the first bearmg has a first centerlme axis and the second bearmg has a second centerlme axis, and the crankshaft cradle has sufficient rigidity to maintain the first and the second centerlme axes within 0 008 mches of one another during operation of the engme at the first engine setting Engmes havmg no more than two main bearmg supports require less precise alignment of the mam beaπngs, because a small amount of beaπng misalignment does not apply a bending moment along the length of the crankshaft (l e , a straight crankshaft axis can, m some cases, e between two miss aligned beaπng supports, but not between three miss aligned beaπng supports) According to the present invention, for engmes having no more than two crankshaft mam bearing supports, the crankshaft cradle has sufficient rigidity to maintain said first and second centerlme axes within 0 040 inches of one another duπng operation of the engme at said first engme setting The engine is considered to have two crankshaft bearing supports if the two bearing supports support more than 85 percent of the crankshaft's radial load Similarly, the crankshaft cradle has sufficient rigidity to liinit rotation of the first bearmg support or eccentric member relative to the second bearmg support or eccentric member about the pivot axis of the cradle to one rotational degree (1°) about pivot axis E at said first engine setting Crankshaft cradles having roller bearings, such as ball beaπngs, also require less precise alignment of the eccentπc mam bearing supports Referring now to Figs 8 and 9. crankshaft cradle 146 has a low rotational inertia, enablmg actuator 138 to rapidly rotate crankshaft cradle 146 about axis E and to rapidly adjust the position of crankshaft centerlme axis A Eccentric section 160 has an outer or bearing diameter 202 that is rotatably housed m a bore 204 m housing 150, and eccentπc section 162 has an outer or bearmg diameter 190 that is rotatably housed m a bore 188 in housmg 150 According to an embodiment of the present invention, to provide a low rotational inertia and a fast response, the ratio of inner diameter 154 to outer diameter 202 is greater than 0 40. and preferably greater than 0 30 Inner bearmg diameter 154 refers to the effective diameter and more specifically the diameter of the hydraulic film separating the crank throw from the journal bearing element 12 For crankshaft cradles roller bearmgs supporting the radial loads of the crankshaft, such as ball, cylmdπcal (mcludmg needle), tapered, and spheπcal roller bearmgs, the effective diameter is the circular path of the individual axes of rotation of the rolling elements In the case of tapered, spheπcal, multiple row bearmgs, and other roller beaπngs havmg a range of circular path diameters, the circular path is measured from the largest circular path of the individual axes of rotation of the rolling elements (Dampers, such as dampers 210 and 212, may be used to dampen deceleration of crankshaft cradle 214. shown m Fig 10 ) Fig 10 shows a partial sectional view of an engme 216 according to the present mvention Fig 10 is similar to Fig 8 except that Fig 10 shows a second hydraulic actuator 218 havmg a piston 220 slidably housed in cylmder 222 Piston 220 is pivotally connected to an arm 224, and arm 224 is pivotaly connected to a crankshaft cradle 214 Cylmder 222 has a flmd lme 226, m flmd commumcation with a first valve 228 and a second valve 230 Second valve 230 may mclude a first check valve 232 Check valve 232 is m flmd communication with a pressuπzed oil feed lme 234 which receives oil under pressure from the oil pump of the engme Cylmder 142 has a flmd lme 236, in flmd commumcation with a third valve 238 and a fourth valve 240 Fourth valve 240 may mclude a second check valve 242 Check valve 242 is m flmd commumcation with a pressuπzed oil feed lme 244 which receives oil under pressure from the oil pump of the engme Oil pass g through valves 228 and 238 returns to the engme sump for eventual recirculation by the pump of the oil pump of the engme Accordmg to the present invention, crankshaft cradle 214 is rotated counterclockwise and crankshaft centerlme axis A is moved towards piston 4 by opemng first valve 228, opemng fourth valve 240. closmg second valve 230, and closmg third valve 238 The position of crankshaft cradle 214 and crankshaft centerlme axis A is retamed in a fixed or near fixed position by closmg first valve 228. leavmg closed second valve 230 (optional), leaving closed valve third valve 238, and leaving open valve 240 Pressuπzed oil flows mto cylmder 142 through feed lme 244, check valve 242, fourth valve 240. and flmd lme 236. causmg crankshaft cradle 214 to rotate counterclockwise and piston 220 to compress oil retained m cylmder 222, where the position of crankshaft cradle 214 becomes fixed or nearly fixed when the pressuπzed oil entering cylinder 142 through feed lme 236 can no longer rotate crankshaft cradle 214 counterclockwise due to the pressure of the oil m cylinder 222. and check valve 242 substantially prevents crankshaft cradle 214 from rotatmg clockwise Accordmg to the present invention crankshaft cradle 214 is rotated clockwise, and crankshaft centerline axis A is moved away from piston 4 by closmg first valve 228, closmg fourth valve 240, opemng second valve 230, and opening third valve 238 The position of crankshaft cradle 214 and crankshaft centerlme axis A is retamed m a fixed or near fixed position as described above, or by leaving closed first valve 228, leavmg closed fourth valve 240, and closmg third valve 238 Those skilled m the art will appreciate that the valve operung and closmg sequences used to adjust and fix the position of crankshaft cradle 214 in engine 216, may be used to adjust the position of crankshaft cradle 112 shown m Fig 7 Other valve operung and closmg sequences may be used to adjust and fix the position of crankshaft cradle 214 in engme 216. and other types of valves may be used to control flow of flmd mto and out of cylmders 142 and 222 accordmg to the present invention The position of crankshaft cradle 214 and crankshaft centerlme axis A may also be retamed in a fixed or near fixed position by closmg first valve 228, opemng second valve 230, closmg valve third valve 238. and operung fourth valve 240 Referring now to Fig 10 (the embodiment shown in Fig 7 may be operated in a similar manner), accordmg to the present mvention feed lmes 244 and 234 are pressurized Preferably standard oil pressure from engme 216 (e g , below 100 psi ) may be used to rotate crankshaft cradle 214 and adjust the position of crankshaft centerlme axis A According to the present mvention, the reversmg loads on crankshaft cradle 214 from the reciprocatmg motion of piston 4 and connectmg arm 6 may be used to rotate crankshaft cradle 214 counterclockwise about axis E and move crankshaft centerlme axis A m a direction generally towards piston 4, and m some embodiments of the present mvention making possible operation of the present invention with small diameter hydraulic pistons 140 and 220, and standard or near standard oil pressure According to the present mvention, the piston and connectmg rod exert forces that change m magnitude on the crankshaft cradle duπng the induction and power strokes of the engme, and the check valve admits and retains flmd m the actuator duπng the induction stroke of the piston, causmg the compression ratio to ratchet up Specifically, crankshaft cradle 214 is rotated counterclockwise about axis E and crankshaft centerlme axis A is moved in a direction generally towards piston 4 by closmg valve 230, opemng valve 228, closmg valve 238, and opemng valve 240 During the mtake or mduction stroke of engme 216, the downward motion of piston 4 and connectmg rod 6 causes crankshaft 152 to exert an upwards force on crankshaft cradle 214, causmg crankshaft cradle 214 to rotate counterclockwise, and flmd to flow out of cylmder 222 through valve 228, and fluid to flow mto cylmder 142 through valve 240 Check valve 242 prevents flmd from leavmg cylmder 142 during the power stroke of piston 4 when the force on crankshaft cradle 214 and crankshaft 152 from piston 4 and connectmg rod 6 reverses and encourages crankshaft cradle 214 to rotate m a clockwise direction about axis E Accordingly, the position of crankshaft centerlme axis A ratchets up m steps, moving in a direction generally towards piston 4 When a desired position of crankshaft centerline axis A is reached, crankshaft cradle 214 may be retamed in position by closmg valve 228 (and optionally opemng valve 230) The embodiments of the present mvention just descπbed work best with engmes where the forces on crankshaft 152 reverse direction, such as in smgle cylinder engmes and some multi cylmder engmes, however, the forces on crankshaft 152 do not have to reverse for the rotational position of the crankshaft cradle to be adjusted according to the present mvention, as the oil pressure m feed lme 244 will encourage crankshaft cradle 214 to rotate m a counterclockwise direction when the pressure in cylmder 142 falls below the pressure m feed lme 244 Preferably, pressure in lme 244 is greater than m cylmder 142 when crankshaft cradle 214 is rotating counterclockwise, to support rotation of crankshaft cradle 214 and to prevent cavitation of oil m cylmder 142 Movmg crankshaft centerlme axis A m a direction generally away from piston 4 may be accomplished by closmg valve 240, operung valve 238, closing valve 228. and opening valve 230 Refemng now to Fig 12, preferably the first hvdraulic piston 264 has the same area as the second piston 266, and flmd from the first hydraulic cylinder is directed directly into the second hydraulic cylinder, thereby preventmg cavitation Fig 14 shows the prefeπed embodunent of the present invention havmg adjustable valve tuning Fig 14 shows an partial view of an engme 300 accordmg to the present invention having a housmg 302, a combustion chamber d, a cylmder bore 20, a cylmder centerlme axis 304, a piston 4 mounted m cylmder 20 for reciprocating motion along cylmder centerlme axis 304, a crankshaft 306 havmg an axis of rotation A and a crank pm 308, mounted m a πdged crankshaft cradle 310, and a connecting rod 6 connectmg piston 4 and crank pm 308 Hydraulic flmd m chambers 312 and 314 acting on surfaces 316 and 318 respectively, rotate crankshaft cradle 310 about an axis E Rotating crankshaft cradle 310 from a first position to a second position causes the crankshaft axis of rotation A to move from centerlme axis F to centerlme axis G, causmg the volume of combustion chamber d to increase and the compression ratio of engme 300 to decrease Accordmg to the present embodiment of the present invention, the crankshaft may be supported m housmg 302 by crankshaft cradle 310 or other functional eccentπc mam beaπng supports rotatably mounted m housing 302 about a eccentπc pivot axis m housmg 302 Engme 300 has one or more pistons 4, one or more mtake camshafts 320, a cylmder head 322, one or more mtake valves 28, one or more intake ports 324, one or more exhaust camshafts 326, one or more exhaust valves 30, and one or more exhaust ports 328 Intake valves 28 and exhaust valves 30 may be opened by direct attack mverted-bucket cam followers (shown) or by other functional means such as finger follower rocker arms (preferably of the roller follower type), centrally pivoted rockers, or another functional arrangement A drive gear 14 (gear teeth contact circle shown, and gear teeth not shown, for all gears and sprockets in Fig 14) is mounted on crankshaft 306 and engages a driven gear 18 mounted on a secondary shaft 330 Dπve gear 14 and driven gear 18 are in mesh and have a mesh direction 332 that pomts generally away from mtake valve 28, and a gear mesh location 334 Gear mesh location 334 is located between eccentric pivot axis E and secondary shaft axis of rotation 336 The crankshaft axis of rotation A is located generally between eccentric pivot axis E and secondary shaft axis of rotation 336 Shaft 330 rotates on axis 336 m housmg 302, and has a pulley, sprocket or other drive means 338 for driving belt or cha 340 Cha 340 rotates a pulley or sprocket or other drive means 342, and sprocket 342 turns intake camshaft 320 The camshaft dπve mcludmg drive means 339, cham 340, and sprocket 342 may be substituted by an alternative functional camshaft dπve The secondary shaft 330 drives the camshaft drive, and the camshaft 320 opens mtake valve 28 More specifically, clockwise rotation of crankshaft 306 rotates gear 14 clockwise, gear 14 then rotates gear 18, shaft 330 and sprocket 338 counterclockwise, sprocket 338 then drives chain 340 generalh counterclockwise Cham 340 then dπves sprocket 342 and camshaft 320 counterclockwise, and camshaft 320 opens intake valve(s) 28 Similarly, gear 14 mounted on crankshaft 306 engages a second dπven gear 344 mounted on a third shaft 346 Shaft 346 rotates on axis 348, and has a pulley, sprocket or other dπve means 350 for driving belt or cham 352 Dπve gear 14 and dπven gear 344 are in mesh and have a mesh direction 354 that points generallv towards mtake valve 28, and a gear mesh location 356 Gear mesh location 356 is located between eccentric pivot axis E and third shaft axis of rotation 348 Crankshaft axis of rotation A is located generally between mesh location 334 and mesh location 356 Mesh direction 354 is generalh in an opposite direction to mesh direction 332 Chain 352 rotates a pulley or sprocket or other drive means 358, and sprocket 358 turns exhaust camshaft 326 The camshaft drive mcludmg dπve means 350, cham 352, and sprocket 358 may be substituted by an alternative functional camshaft drive Clockwise rotation of crankshaft 306 rotates gear 14 clockwise, gear 14 then rotates gear 344, shaft 346 and sprocket 350 counterclockwise, sprocket 350 then drives cham 352 generally counterclockwise Chain 352 then drives sprocket 342 and camshaft 326 counterclockwise, and camshaft 326 opens exhaust valve(s) 30 According to the present invention, exhaust valve 30 may be dπven by shaft 330 and or the mtake camshaft dπve, where the exhaust and mtake cam shafts are phase shifted m the same direction when compression ratio is changed In the embodiment shown in Fig 15b. cylmder 20 (shown m Fig 14) has two mtake valves, intake camshaft 320 and exhaust camshaft 326 are dπven by shaft 330 and the intake camshaft drive, and, intake camshaft 320 has a phase shifter 360 for adjustmg the phase tuning between the two mtake valves, thereby providing a low friction valve tram with adjustable valve control for providing low engine pumping losses Alternatively, mtake camshaft 320 and exhaust camshaft 326 may be driven by shaft 346 and the exhaust camshaft drive, where the exhaust and mtake camshafts are phase shifted m the same direction when compression ratio is changed According to the present invention, the timing of exhaust valve closmg (EVC) and the timing of intake valve opemng (IVO) is adjusted to prevent valves 28 and 30 from striking piston 4 and unproved idle stability (and in particular when crankshaft 306 is located on axis F and engme 300 is operatmg at its maxunum compression ratio settmg), and to provide improved flow of exhaust out of chamber d and mto exhaust port 328, and improved flow of intake air through port 324 and mto chamber d (and in particular when crankshaft 306 is located on axis G and engme 300 is operating at its lniniinum compression ratio settmg) Accordmg to the present mvention, the period of time that valves 28 and 30 are both open, the valve overlap period, is adjusted by rotatmg ridged crankshaft cradle 310 about axis E Rotating ridged crankshaft cradle 310 about axis E from its first position (closest to mtake valve 28) to its second position causes the axis of rotation of crankshaft 306 to move from axis location F to axis location G, and the phase relationship between gear 14 and gear 18 to adjust Specifically, rotatmg πdged crankshaft cradle 310 about axis E from its first position to its second position causes gear 18 to rotate counterclockwise, and camshaft 320 to open valve 28 earlier relative to the tuning of mtake valve opemng when crankshaft 306 is located on axis F More specifically, intake valve 28 has a later timing of opemng relative to crankshaft 306 at the first crankshaft position F than at the second crankshaft axis position G Similarly, rotating πdged crankshaft cradle 310 about axis E from its first position to its second position causes the axis of rotation of crankshaft 306 to move from axis location F to axis location G. and the phase relationship between gear 14 and gear 344 to adjust Specifically, rotatmg ridged crankshaft cradle 310 about axis E from its first position to its second position causes gear 344 to rotate clockwise, and camshaft 326 to close valve 30 later relative to the tuumg of exhaust valve closing when crankshaft 306 is located on axis F Accordingly . the period of time that mtake valve 28 and exhaust valve 30 are both open, the valve overlap period (VOL), is greater when crankshaft 306 is located on axis G than when crankshaft 306 is located on axis F The change m phase between gear 14 and gear 18 from the first crankshaft position to the second crankshaft position is, among other factors, a function of the distance between axis F and axis G, and the distance between axis A and axis 336 Sumlarly, the change in phase between gear 14 and gear 344 from the first crankshaft position to the second crankshaft position is. among other factors, a function of the distance between axis F and axis G, and the distance between axis A and axis 348 Accordmg to the present mvention, the magmtude of phase change of gear 18 can be the same or different than the magmtude of phase change of gear 344 In the embodiment of the present mvention shown m Fig 14, the centerlme distance between axis 348 and A is shorter than the centerlme distance between axis 336 and A, and accordingly the magmtude of phase change is greater for gear 344 than gear 18 from the first crankshaft position to the second crankshaft position In the embodiment of the present mvention shown in Fig 14, the distance between gear 14 and 18 slightly changes as axis A moves from position F to position G Similarly, the distance between gear 14 and 344 slightly changes as axis A moves from position F to position G Accordmg to the present mvention, engme 300 has power output means havmg a variable distance between gear 14 and gear 18, and gear 14 and gear 344, and according to the present mvention moving the crankshaft centerlme axis from a first position to a second position adjusts the phase of exhaust cam 326, the phase of mtake cam 320, and/or the period of time that both mtake and exhaust valves are open Those skilled in the art will appreciate that the phase of exhaust cam 326, the phase of mtake cam 320, and/or the period of tune that both mtake and exhaust valves are open may be adjusted according to the present mvention with other power output coupling means such as shown m Fig 4 of Deutsches Patentant DE 36 44721 Al, dated December 30. 1986 and July 14, 1988 Those skilled m the art will appreciate that cham 340 and or 352 may be replaced with one or more gears that drive the cam lobes, and that the phase change of gear 18 and or 344 according to the present invention is unaffected Those skilled m the art will appreciate that accordmg to the present invention both camshafts may be dπven by a smgle cham or belt (e g , 340 or 352) or gear set, with the phase change caused by movmg the centerlme axis of crankshaft 306 providing some benefit to at least one of the cam shafts A phase adjuster may be employed to adjust the phase relationship between the two camshafts A control system mav be employed that prevents crankshaft 306 from bemg raised from position G to position F until after one or more phase shifters have adjusted the phase relation ship of one or both (or more) camshafts to prevent valves 28 and 30 from striking piston 4 Fig 15 shows Exhaust camshaft 326 and intake camshaft 320 along cut lines S15-S15, shown in Fig 14 Exhaust camshaft 326 has cam lobes 382 and 384 Those skilled in the art will appreciate that exhaust camshaft 326 can have one or more cam lobes Intake camshaft 320 has cam lobes 386 and 388 Those skilled m the art will appreciate that mtake camshaft 320 can have one or more cam lobes In the embodiment of the present invention shown m Fig 15, cam shaft 320 includes a prmiary drive shaft 390 and a follower 392, and an optional phase shifter 360 for changing the phase relationship between cam lobe 386 and 388 Fig 17 shows crankshaft 306. shaft 330, and ridged crankshaft cradle 310 along cut lmes S17-S17 shown in Fig 14 In the embodiment of the present invention shown m Figs 14 and 17. shaft 330 serves as a balance shaft According to the present mvention, engme 300 is balanced by niinirnizing the distance between crankshaft axis A and balance shaft axis 336, and by locatmg the cylinder centerlme axis 304 between crankshaft axis A and balance shaft axis 336 Crankshaft 306 has a rotational speed and a rotational direction, and balance shaft 330 has a rotational speed and a rotational direction Balance shaft 330 has the same rotational speed as crankshaft 306, and balance shaft 330 has an opposite rotational direction to crankshaft 306 for primary balancing Balance shaft 330 has a bow 394 that bows inwardly across beaπng diameter 400 and the centerlme 336 of balance shaft 330, and that provides clearance for rod 6 duπng rotation of crankshaft 306 and balance shaft 330 Balance shaft 330 and crankshaft 306 rotate at the same speed but m opposite directions Balance shaft 330 has bearmgs 396 and 398 each havmg an imaginary projected beaπng diameter cylmder 400 runnmg parallel to axis 336 (the bearmg diameter cylmders of bearmgs 396 and 398 are dashed in as an imaginary lmes 400 m Fig 17) For engmes having anti-friction bearmgs such as roller or ball bearings, the bearing diameter cylmder 400 is measured from the inner raceway According to the present mvention, rod 6 crosses at least one beaπng diameter cylinder 400 and preferably rod 6 crosses balance shaft centerlme axis 336 during rotation Balance shaft 330 has bows 402 that bow m generally the opposite direction of bow 394, and provide clearance for counter weights 404 Preferably, bow 402 inwardly crosses at least one bearmg diameter cylinder 400 to provide clearance for counterweights 404 bow 402 may inwardly cross balance shaft axis 336 (shown in Fig 17b) Preferably, counterweights 404 cross at least one bearmg diameter cylmder 400 duπng rotation of crankshaft 306, and counterweight 404 may cross balance shaft axis 336 (shown in Fig 17b) Counter weights 404 have a radius 406 to closely pass clear of balance shaft 330, and balance shaft 330 has radiuses 408 to closely pass clear of counterweights 404 The outwardly force of counterweights 404 may be mcreased by adding a heavy metal to counterweights 404 such as tungsten or lead, or by increasmg the length 410 of the crankshaft In the embodiment of the present mvention shown m Fig 17, length 410 is greater than 90 percent of the radius r of bore 20, and preferably greater than the radius of bore 20, and bow 394 inwardly crosses balance shaft centerlme 336 Distance 412 between crankshaft axis A and balance shaft axis 336 is shorter than the length of the stroke of piston 4, and preferably the distance between crankshaft axis A and balance shaft axis 336 is shorter than 90 percent of the length of the stroke of piston 4, thereby providing a reduced balance shaft spacmg and improved engme balancing Refemng now to Figs 14 and 17, crankshaft axis A and balance shaft axis 336 are separated by a distance 412 having a midpoint 414 Crankshaft axis of rotation A and midpomt 414 are separated by a distance 416, bemg half the length of distance 412 Cylmder centerline axis 304 and crankshaft axis A are separated by a distance 418 Cylmder centerlme axis 304 and nndpomt 414 are separated by a distance 420 Distances 412, 416, 418, and 420 change in length a very small amount with change in compression ratio. The change in length of distance 412, 416, 418, and 420 may be ignored with respect to engine balancing. Cylinder centerline axis 304 passes between crankshaft axis A and balance shaft axis 336. Engine 300 has a balance off-set ratio of distance 420 to distance 416 of no more than 0.90. In detail, the distance between the cylinder centerline axis 304 and midpoint 414 is at least 90 percent of the length between the crankshaft axis A and midpoint 414, thereby providing improved primary balance, and in particular providing improved primary balance in engines having only one balance shaft rotating at crankshaft rotational speed. Preferably, length 418 is greater than 20% of the distance between crankshaft axis A and balance shaft axis 336, and preferably length 418 is greater than 25% of the distance between crankshaft axis A and balance shaft axis 336. At a minimum, length 418 is greater than 15% of the distance between crankshaft axis A and balance shaft axis 336. Crank pin 308 has an axis B. The stroke of piston 4 is approximately equal to twice the distance from crank pin axis B to crankshaft axis A. Engine 300 has a cylinder off-set ratio (or "off set ratio") of length 418 to the stroke of piston 4. Ideally, distance 418 between crankshaft axis A and cylinder centerline axis 304 is at least 10 percent as long as the length of the stroke of piston 4, thereby providing a reduced balance off-set and improved engine balancing. Preferably, said off-set ratio is at least 0.03, therby providing improved primary or secondary engine balancing. According to the present invention, the power stroke of engine 300 drives the big end of rod 6 down between the crankshaft axis A and the balance shaft axis 336, and more specifically, the mesh direction 332 between gear 14 and gear 18 points generally away from piston 4, thereby providing reduced friction in addition to improved balance. The present invention significantly improved engine balancing, and in particular for engines having fewer than three pistons where primary balancing is poor. Counterweight 422 is mounted on crankshaft 306, and counterweight 424 is mounted on shaft 330. Preferably the polar moment of inertia of 422 is the same or almost the same as the polar moment of inertia of 424. Counterweight 422 is mounted on the front end of engine 300 and crosses axis 336, and counterweight 424 is mounted on the back end of engine 300 and crosses axis A. Counterweight 424 is located on the same end of shaft 330 as gear 18 and the power output of engine 300 is through the same end of shaft 330, and power may be taken out through shaft end 426, through gear 18, or through other suitable means. Crankshaft 306 is sufficiently ridged to prevent unacceptable levels of vibration, and in particular harmonic vibration between flywheel 422 and flywheel 424. Those skilled in the art will appreciate that engine 300 may have other arrangements according to the present invention. Refeπing now to Figs. 14, 14b, and 14c, crankshaft 306 is moved by crankshaft cradle 310, or another type of eccentric main bearing supports, towards cylinder head 322 of engine 300 during a portion of the power stroke of piston 4. Fig. 14b shows a free-body diagram of forces acting on crankshaft 306 located at 90 crank angle degrees clockwise from top dead center as shown in Fig. 14. In more detail, Fig. 14b shows piston gas force acting through crank pin axis B (located at 90 crank angle degrees clockwise after top dead), gear torque acting on crankshaft 306 at gear mesh location 334, and a reaction force acting on crankshaft cradle 310 at crankshaft rotational axis A Fig 14c shows a detailed view of a ratchet 444 Ratchet 444 attaches to engme 300 shown in Fig 14 Referπng now to Figs 14 and 14b, according to the present mvention, crankshaft 306 serves as a lever, and gear mesh 334 serves as a fulcrum Crank pm 308 is located at a first end of the "lever" (e g , crankshaft), and the crankshaft mam bearmgs 395 located about crankshaft axis A (shown m Fig 17) are located at the other end of the "lever", with gear mesh 334 located between the first and second ends of the "lever" and serving as a fulcrum The force of the power stroke (at approximately 90 crank angle degrees after top dead center) beaπng down on the crank pm (the first end of the lever) causes an upward force on crankshaft cradle 310 by crankshaft mam beaπngs 395 (e g , located at the second end of the lever), where gear mesh 334 acts as a fulcrum A ratchet 444 permits crankshaft cradle 310 to rotate (counterclockwise as shown m Fig 14) causmg the centerlme of crankshaft 306 to move towards cylmder head 322 Ratchet 444 prevents crankshaft cradle 310 from rotating m the opposite direction, thereby causmg crankshaft 306 to advance towards cylmder head 322 m steps In detail, crank pm centerlme B has an orbital diameter 446, and drive gear 14 has a pitch diameter 448 Crankshaft 306 is moved towards cylmder head 322 during a portion of the power stroke of piston 4 by placing the orbital diameter 446 of crank pm 308 outside of the pitch diameter 448 of the dπve gear 14, placing dπve gear 14 m mesh with driven gear 18, plac g crank pm 308 during the power stroke of piston 4 on the opposite side of gear mesh location 334 from crankshaft axis A and crankshaft mam beaπngs 395, firmg the engme. and ratcheting crankshaft cradle 310 and crankshaft 306 towards cylinder head 322 m steps, where the crankshaft mam bearmgs 395 and the crankshaft 306 pivot toward the cylmder head about gear mesh 334 under the force away from the cylmder head 322 of the power stroke acting on crank pm 308 According to the present invention, the force of the piston on crank pin 308 during a portion of the power stroke may not be sufficient to cause the crankshaft to move towards cylmder head 322. or move the crankshaft to cylmder head 322 quickly enough According to the present mvention, oil pressure m chamber 312 may be sufficiently mcreased to cause crankshaft cradle 310 to rotate and crankshaft 306 to move towards cylmder head 322 under the combined force of the oil pressure m chamber 312 and the force of the piston on crank pm 308 during a portion of the power stroke Fig 14c shows ratchet 444 According to the present mvention, crankshaft cradle 310 may be moved m steps by a hydraulic ratchet (shown m Fig 14c), a mechanical ratchet, an electrical ratchet, a hydro-mechanical, electπc ratchet, or another type of functional ratchet Fig 14c shows a hydraulic ratchet that is similar to the hydraulic system shown m Fig 9, except that the outflow from chamber 314 is ducted into the inflow of chamber 312. and the outflow of chamber 312 is ducted mto the inflow of chamber 314 thereby reducing actuator power and preventing hydraulic cavitation Referring to Fig 14c, crankshaft cradle 310 is moved clockwise for mo ing crankshaft 306 away from cylinder head 322 by closmg valves 240 and 450 and opemng valve 230 Clockwise motion of crankshaft cradle 310 caused be forces on crankshaft 306 causes fl d to be forced out of chamber 312 into duct 236 mto duct 234. through open valve 230, through one-way valve 232 through duct 226, and mto chamber 314 Reverse flow is prevented by one-way or check valve 232 Crankshaft cradle 310 is moved counterclockwise for movmg crankshaft 306 towards cylmder head 322 by operung valve 240 and closmg valve 230 Counterclockwise motion of crankshaft cradle 310 caused be forces on crankshaft 306 causes flmd to be forced out of chamber 314 mto duct 226, mto duct 244, through open valve 240, through one-way valve 242 through duct 236, and mto chamber 312 Reverse flow is prevented by one-way or check valve 242 Counterclockwise motion of crankshaft cradle 310 and movement of crankshaft 306 towards cylmder head may be assisted by opemng valve 450 Opemng valve 450 permits feed oil under pressure to enter chamber 312 through valve 460 and duct 236, causmg crankshaft cradle 310 to rotate counterclockwise and flmd m chamber 314 to be forced through duct 226, though duct 452, through valve 450, through one way valve 454, and through dram pipe 456 mto an oil pan 458 or mto another functional drainage receptacle Drain pipe 456 opens mto oil pan 458 below the surface of the oil m the pan in order to prevent air from entering oil feed lmes 456, 452, 226 and ultimately chambers 314 and 312 Valves 230 and 240 may be located on the same spool and opened together In systems were actuator 444 is operated with valve 450 closed, the volume displaced from chamber 314 must be the same as the volume added to chamber 312 for a given amount of rotation of crankshaft cradle 310 Piston type chambers, or another functional type of hydraulic chambers, may be used as an alternative to chambers 312 and 314

Claims: CLAIMS 1 A vanable compression ratio engme with an adjustable valve timing having an engine housmg and a crankshaft havmg a crankshaft axis of rotation, a first crankshaft axis position relative to said housmg, and a second crankshaft axis position relative to said housmg, and one or more eccentπc mam bearing supports, rotatably mounted in said housmg about an eccentπc pivot axis in said housmg, for adjusting said crankshaft axis from said first crankshaft axis position to said second crankshaft axis position, a dπve gear mounted on said crankshaft, and a dπven gear mounted on a secondary shaft having a second shaft axis of rotation, said second shaft axis of rotation bemg fixed m said housmg, said drive gear being in mesh with said dπven gear, said dπve gear and said dπven gear havmg a first mesh direction and a first mesh location, a first mtake valve and a camshaft havmg a camshaft dπve, said secondary shaft dπves said camshaft drive, and said camshaft opens said mtake valve. wherem said first mesh direction pomts generally away from said first intake valve, said first crankshaft axis position bemg closer to said first valve than said second crankshaft axis position, said camshaft has a first phase timing relative to said crankshaft at said first crankshaft axis position, and said camshaft has a second phase timing relative to said crankshaft at said second crankshaft axis position, wherein said mtake valve has a later timing of opening relative to said crankshaft at said first crankshaft axis position than at said second crankshaft axis position 2 The vaπable compression ratio engine with an adjustable valve timing of claim 1 , wherem said dπve gear and said dπven gear have a gear mesh location, said gear mesh location being between said eccentπc pivot axis and said second shaft axis of rotation 3 The variable compression ratio engine with an adjustable valve timing of claim 2, wherem said crankshaft axis of rotation is located generally between said eccentric pivot axis and said second shaft axis of rotation 4 The vaπable compression ratio engine with an adjustable valve tuning of claim 2, wherein said eccentric pivot axis is located between said crankshaft axis of rotation and said second shaft axis of rotation 5 The variable compression ratio engine with an adjustable valve timing of claim 1 further havmg a first exhaust valve and an exhaust valve dπve, wherein said first exhaust valve and said first mtake valve have a first valve overlap peπod at said first crankshaft axis position and a second valve overlap peπod at said second crankshaft axis position, said first valve overlap peπod bemg shorter than said second valve overlap peπod 6 The vaπable compression ratio engme with an adjustable valve timing of claim 5, further havmg a second dπven gear mounted on a third shaft having a third axis of rotation, said third axis of rotation bemg fixed m said housmg, said second dπven gear bemg m mesh with a gear mounted on said crankshaft and having a second mesh direction and a second mesh location, said crankshaft axis of rotation bemg located generally between said first mesh location and said second mesh location, and said second mesh direction bemg generally opposite to said first mesh direction 7 The variable compression ratio engme with an adjustable valve timing of claim 1 , further havmg a first exhaust valve, said exhaust valve being dπven by said dπven gear mounted on said secondary shaft 8 The vaπable compression ratio engine with an adjustable valve timing of claun 7, further havmg a second mtake valve and a intake valve tuning phase shifter for adjustmg the tuning of said second mtake valve relative to said first mtake valve 9 An internal combustion engme havmg a crankshaft, said crankshaft havmg a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connectmg rod connecting said piston and said crankshaft, a cylinder havmg a cylmder centerlme axis, said piston bemg mounted in said cylmder for reciprocating motion along said cylmder centerlme axis, a balance shaft havmg a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction bemg opposite, said balance shaft rotational speed and said crankshaft rotational speed bemg the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation havmg a mid pomt, a second distance from said crankshaft axis of rotation to said mid pomt, said second distance bemg equal m length to half of said first distance, and a third distance from said cylinder centerlme axis to said mid pomt, wherem said cylmder centerlme axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherem said engme has a balance off-set ratio of said third distance to said second distance of no more than 0 90, thereby providmg an unproved primary balance 10 The internal combustion engme of claun 9, wherem said reciprocating piston machine has only one of said balance shafts rotatmg at said crankshaft rotational speed 11 The internal combustion engme of claun 9, wherem said cylinder centerlme axis and said crankshaft axis of rotation are separated by a fourth distance, said fourth ώstance bemg at least 15 percent as long as said first distance 12 The internal combustion engme of claun 9, wherem said piston has a stroke havmg a length, said cylmder centerlme axis and said crankshaft axis of rotation are separated by a fourth distance, said fourth distance bemg at least 10 percent as long as said stroke 13 The internal combustion engme of claun 9, wherem said first distance is less than the length of said stroke 14 The internal combustion engme of claun 9, further having a drive gear mounted on said crankshaft and a dπven gear mounted on said balance shaft, said drive gear bemg m mesh with said dπven gear, said dπve gear and said dπven gear havmg a mesh direction, wherem said mesh direction pomts generally away from said piston, thereby providing improved balance and reduced friction 15 The internal combustion engme of claim 9, wherein said engme has fewer than three of said pistons 16 The internal combustion engme of claun 14, wherein said engine is a vanable compression ratio engine 17 An internal combustion engme having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed, a crankshaft rotational direction, and a stroke, said stroke having a stroke length, a piston, and a connectmg rod connectmg said piston and said crankshaft, a balance shaft havmg a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, a cylmder having a cylmder centerline axis, said piston bemg mounted in said cylmder for reciprocating motion along said cylmder centerlme axis, said c> lmder centerlme axis bemg located between said crankshaft axis of rotation and said balance shaft axis of rotation said balance shaft rotational direction and said crankshaft rotational direction bemg opposite, said cylmder centerlme axis and said crankshaft axis of rotation bemg separated by a first distance. wherem said internal combustion engme has a cylmder off-set ratio of said first distance to said stroke length, said off-set ratio bemg at least 0 03, thereby providing unproved primary balancing 18 A method for moving a crankshaft of an engme about a pivot axis towards the cylinder head of the engme dur g a portion of the power stroke of a first piston for adjusting the compression ratio of the engme, wherem the engme has a crankshaft havmg a crankshaft axis of rotation, crankshaft mam beaπngs, and a crank pm having an orbital diameter, a connecting rod for connecting said first piston and said crank pin, an engme housmg and one or more eccentπc mam beaπng supports rotatably mounted m said housmg about a pivot axis in said housing for pivotaly supporting said crankshaft m said crankshaft mam bearmgs, a drive gear mounted on said crankshaft for transferπng power from said crankshaft to a power takeoff shaft, said drive gear havmg a pitch diameter, and a dπven gear mounted on said power take-off shaft, and, a first ratchet for ratcheting said eccentπc mam beaπng supports m a first pivot direction, said first ratchet bemg disengagable for rotation of said eccentπc mam beaπng supports in a second pivot direction, compπsmg the steps of, placmg the orbital diameter of the crank pm outside of the pitch diameter of the dπve gear, placmg the drive gear m mesh with the driven gear, placmg the crank pm durmg the power stroke of the first piston on the opposite side of the gear mesh from the crankshaft mam bearings, firing the engme, and ratchetmg the crankshaft towards the cylinder head m steps, wherem the crankshaft mam bearings pivot toward the cylinder head about the gear mesh under the force away from the cylmder head of the power stroke actmg on the crank pin, and the crankshaft is ratcheted towards the cylmder head m steps AMENDED CLAIMS [received by the International Bureau on 28 December 2000 (28.12.00); original claims 1, 4, 9, 14 and 15 cancelled; original claims 2, 5, 7, 10-13, 16 and 17 amended; new claim 19 added: remaining claims unchanged (7 pages)] Claim 1 is cancelled A vaπable compression ratio engme with an adjustable valve timing having an engine housmg and a crankshaft having a crankshaft axis of rotation. a first crankshaft axis position relative to said housmg, and a second crankshaft axis position relative to said housmg, and one or more eccentπc mam beaπng supports, rotatably mounted in said housing about an eccentπc pivot axis m said housing, for adjustmg said crankshaft axis from said first crankshaft axis position to said second crankshaft axis position. a dnve gear mounted on said crankshaft, and a dπven gear mounted on a secondarv shaft having a second shaft axis of rotation, said second shaft axis of rotation bemg fixed in said housing, said dπve gear bemg in mesh with said dπven gear, said dπve gear and said dπven gear havmg a first mesh direction and a first mesh location. a first intake valve and a camshaft having a camshaft dπve. said secondary shaft dπves said camshaft dπve. and said camshaft opens said mtake valve. wherem said first mesh direction pomts generally away from said first mtake valve, said first crankshaft axis position bemg closer to said first valve than said second crankshaft axis position, said camshaft has a first phase tuning relative to said crankshaft at said first crankshaft axis position, and said camshaft has a second phase tuning relative to said crankshaft at said second crankshaft axis position. wherem said mtake valve has a later timing of opening relative to said crankshaft at said first crankshaft axis position than at said second crankshaft axis position. wherem said first mesh location is located between said eccentπc pivot axis and said second shaft axis of rotation 3 The vaπable compression ratio engine with an adjustable valve timing of claim 2. wherem said crankshaft axis of rotation is located generally between said eccentπc pivot axis and said second shaft axis of rotation 4 Claim 4 is cancelled. 5 A vaπable compression ratio engine with an adjustable valve timing having an engine housing, a crankshaft having a crankshaft axis of rotation, a first exhaust valve and an exhaust valve dπve. a first crankshaft axis position relative to said housing, and a second crankshaft axis position relative to said housing, and 33 one or more eccentric main bearing supports, rotatably mounted in said housing about an eccentric pivot axis in said housing, for adjusting said crankshaft axis from said first crankshaft axis position to said second crankshaft axis position, a drive gear mounted on said crankshaft, and a driven gear mounted on a secondary shaft having a second shaft axis of rotation, said second shaft axis of rotation being fixed in said housing, said drive gear being in mesh with said driven gear, said drive gear and said driven gear having a first mesh direction and a first mesh location, a first intake valve and a camshaft having a camshaft drive, said secondary shaft drives said camshaft drive, and said camshaft opens said intake valve, wherein said first mesh direction points generally away from said first intake valve, said first crankshaft axis position being closer to said first valve than said second crankshaft axis position, said camshaft has a first phase timing relative to said crankshaft at said first crankshaft axis position, and said camshaft has a second phase timing relative to said crankshaft at said second crankshaft axis position, wherein said intake valve has a later timing of opening relative to said crankshaft at said first crankshaft axis position than at said second crankshaft axis position, wherein said first exhaust valve and said first intake valve have a first valve overlap period at said first crankshaft axis position and a second valve overlap period at said second crankshaft axis position, said first valve overlap period being shorter than said second valve overlap period. 6. The variable compression ratio engine with an adjustable valve timing of claim 5, further having a second driven gear mounted on a third shaft having a third axis of rotation, said third axis of rotation being fixed in said housing, said second driven gear being in mesh with a gear mounted on said crankshaft and having a second mesh direction and a second mesh location, said crankshaft axis of rotation being located generally between said first mesh location and said second mesh location, and said second mesh direction being generally opposite to said first mesh direction. 7. A variable compression ratio engine with an adjustable valve timing having an engine housing, a crankshaft having a crankshaft axis of rotation, and a first exhaust valve, a first crankshaft axis position relative to said housing, and a second crankshaft axis position relative to said housing, and one or more eccentric main bearing supports, rotatably mounted in said housing about an eccentric pivot axis in said housing, for adjusting said crankshaft axis from said first crankshaft axis position to said second crankshaft axis position, a drive gear mounted on said crankshaft, and a driven gear mounted on a secondary shaft having a second shaft axis of rotation, said second shaft axis of rotation being fixed in said housing, said drive gear being in mesh with said driven gear, said drive gear and said driven gear having a first mesh direction and a first mesh location, a first intake valve and a camshaft having a camshaft drive, said secondary shaft drives said camshaft drive, and said camshaft opens said intake valve, wherein said first mesh direction points generally away from said first intake valve, said first crankshaft axis position being closer to said first valve than said second crankshaft axis position, said camshaft has a first phase timing relative to said crankshaft at said first crankshaft axis position, and said camshaft has a second phase timing relative to said crankshaft at said second crankshaft axis position, wherein said intake valve has a later timing of opening relative to said crankshaft at said first crankshaft axis position than at said second crankshaft axis position, and said exhaust valve is driven by said driven gear mounted on said secondary shaft. 8. The variable compression ratio engine with an adjustable valve timing of claim 7, further having a second intake valve and a intake valve timing phase shifter for adjusting the timing of said second intake valve relative to said first intake valve. 9. Claim 9 is cancelled. 10. An internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connecting rod connecting said piston and said crankshaft, a cylinder having a cylinder centerline axis, said piston being mounted in said cylinder for reciprocating motion along said cylinder centerline axis, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation having a mid point, a second distance from said crankshaft axis of rotation to said mid point, said second distance being equal in length to half of said first distance, and a third distance from said cylinder centerline axis to said mid point, wherein said engine has only one of said balance shafts rotating at said crankshaft rotational speed, wherein said cylinder centerline axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherein said engine has a balance off-set ratio of said third distance to said second distance of no more than 0.90, thereby providing an improved primary balance. 11. An internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connecting rod connecting said piston and said crankshaft, a cylinder having a cylinder centerline axis, said piston being mounted in said cylinder for reciprocating motion along said cylinder centerline axis, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation having a mid point, a second distance from said crankshaft axis of rotation to said mid point, said second distance being equal in length to half of said first distance, and a third distance from said cylinder centerline axis to said mid point, wherein said cylinder centerline axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherein said engine has a balance off-set ratio of said third distance to said second distance of no more than 0.90, wherein said cylinder centerline axis and said crankshaft axis of rotation are separated by a fourth distance, said fourth distance being at least 15 percent as long as said first distance, thereby providing an improved primary balance. 12. An internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connecting rod connecting said piston and said crankshaft, a cylinder having a cylinder centerline axis, said piston being mounted in said cylinder for reciprocating motion along said cylinder centerline axis, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation having a mid point, a second distance from said crankshaft axis of rotation to said mid point, said second distance being equal in length to half of said first distance, and a third distance from said cylinder centerline axis to said mid point, said cylinder centerline axis and said crankshaft axis of rotation are separated by a fourth distance, wherein said cylinder centerline axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherein said engine has a balance off-set ratio of said third distance to said second distance of no more than 0.90, wherein said piston has a stroke having a length, said fourth distance being at least 10 percent as long as said stroke, thereby providing an improved primary balance. 13. An internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connecting rod connecting said piston and said crankshaft, a cylinder having a cylinder centerline axis, said piston being mounted in said cylinder for reciprocating motion along said cylinder centerline axis, and said piston has a stroke having a length, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation having a mid point, a second distance from said crankshaft axis of rotation to said mid point, said second distance being equal in length to half of said first distance, and a third distance from said cylinder centerline axis to said mid point, wherein said cylinder centerline axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherein said engine has a balance off-set ratio of said third distance to said second distance of no more than 0.90, and said first distance is less than the length of said stroke, thereby providing an improved primary balance. 14. Claim 14 is cancelled. 15. Claim 15 is cancelled. 16. A variable compression ratio internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed and a crankshaft rotational direction, a piston, and a connecting rod connecting said piston and said crankshaft, a cylinder having a cylinder centerline axis, said piston being mounted in said cylinder for reciprocating motion along said cylinder centerline axis, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, a first distance from said crankshaft axis of rotation to said balance shaft axis of rotation having a mid point, a second distance from said crankshaft axis of rotation to said mid point, said second distance being equal in length to half of said first distance, and a third distance from said cylinder centerline axis to said mid point, wherein said cylinder centerline axis passes between said crankshaft axis of rotation and said balance shaft axis of rotation, wherein said engine has a balance off-set ratio of said third distance to said second distance of no more than 0.90, thereby providing an improved primary balance, wherein said engine has a drive gear mounted on said crankshaft and a driven gear mounted on said balance shaft, said drive gear being in mesh with said driven gear, said drive gear and said driven gear having a mesh direction, wherein said mesh direction points generally away from said piston, thereby providing improved balance and reduced friction. 17. An internal combustion engine having a crankshaft, said crankshaft having a crankshaft axis of rotation, a crankshaft rotational speed, a crankshaft rotational direction, and a stroke, said stroke having a stroke length, a piston, and a connecting rod connecting said piston and said crankshaft, a balance shaft having a balance shaft axis of rotation, a balance shaft rotational speed and a balance shaft rotational direction, a cylinder having a cylinder centerline axis, said cylinder centerline axis being located between said crankshaft axis of rotation and said balance shaft axis of rotation, said balance shaft rotational direction and said crankshaft rotational direction being opposite, said balance shaft rotational speed and said crankshaft rotational speed being the same, said cylinder centerline axis and said crankshaft axis of rotation being separated by a first distance, wherein said engine has only one of said balance shafts rotating at said crankshaft rotational speed, wherein said internal combustion engine has a cylinder off-set ratio of said first distance to said stroke length, said off-set ratio being at least 0.03, thereby pro * viding improved primary balancing. 18. A method of moving a crankshaft of an engine about a pivot axis towards the cylinder head of the engine during a portion of the power stroke of a first piston for adjusting the compression ratio of the engine, wherein the engine has a crankshaft having a crankshaft axis of rotation, crankshaft main bearings, and a crank pin having an orbital diameter, a connecting rod for connecting said first piston and said crank pin, an engine housing and one or more eccentric main bearing supports rotatably mounted in said housing about a pivot axis in said housing for pivotaly supporting said crankshaft in said crankshaft main bearings, a drive gear mounted on said crankshaft for transferring power from said crankshaft to a power takeoff shaft, said drive gear having a pitch diameter, and a driven gear mounted on said power take-off shaft, and, a first ratchet for ratcheting said eccentric main bearing supports in a first pivot direction, said first ratchet being disengagable for rotation of said eccentric main bearing supports in a second pivot direction, comprising the steps of; placing the orbital diameter of the crank pin outside of the pitch diameter of the drive gear; placing the drive gear in mesh with the driven gear; placing the crank pin during the power stroke of the first piston on the opposite side of the gear mesh from the crankshaft main bearings; firing the engine; and ratcheting the crankshaft towards the cylinder head in steps, wherein the crankshaft main bearings pivot toward the cylinder head about the gear mesh under the force away from the cylinder head of the power stroke acting on the crank pin, and the crankshaft is ratcheted towards the cylinder head in steps. 19. A method of moving a crankshaft of an engine about a pivot axis towards the cylinder head of the engine for adjusting the compression ratio of the engine, wherein the engine has a crankshaft having a crankshaft axis of rotation, crankshaft main bearings, and a crank pin having an orbital diameter, a connecting rod for connecting said first piston and said crank pin, an engine housing and one or more eccentric main bearing supports rotatably mounted in said housing about a pivot axis in said housing for pivotaly supporting said crankshaft in said crankshaft main bearings, a drive gear mounted on said crankshaft for transferring power from said crankshaft to a power takeoff shaft, and a driven gear mounted on said power take-off shaft, a hydraulic ratchet having hydraulic fluid for ratcheting said eccentric main bearing supports in a first pivot direction, said hydraulic ratchet having a pressurized hydraulic feed line having a check valve and having a first feed line valve, comprising the steps of; placing the drive gear in mesh with the driven gear; firing the engine; owning the first feed line valve; trapping hydraulic fluid in the hydraulic ratchet with the check valve; and ratcheting the crankshaft towards the cylinder head in steps, wherein the check valve traps hydraulic fluid in the hydraulic ratchet and the crankshaft is ratcheted towards the cylinder head in steps.

Inventors: MENDLER EDWARD CHARLES, MENDLER, EDWARD, CHARLES
Applicants: MENDLER EDWARD CHARLES, MENDLER, EDWARD, CHARLES
Application year: 2000
Main classification: F02B75
IPC codes: F02B75, F02F7