BACKGROUND OF THE INVENTION The useful work of internal combustion engines is limited by their mechanical efficiency.

On average, about 85 % of the work available on the piston at full load is available as useful work on the flywheel due to internal engine friction. At lower loads, the above figure is even lower. Piston assembly friction (piston and piston rings) alone can account for up to 75 % of the overall mechanical losses. Thus, piston friction reduction is highly desirable. Furthermore, engine parts are subject to wear that eventually limits the engine power and efeciency as well as increasing oil consumption, and increasing exhaust emissions. The most critical wear occurs on the piston rings and cylinder liners.

Excessive wear requires engine overhaul or replacement. Thus, liner and ring wear reduction is also highly desirable.

Piston rings have two primary functions: Limiting oil flow into the combustion chamber and minimizing blowby (leak of high pressure combustion gas from the combustion chamber into the crankcase). Both functions are accomplished as high pressure combustion gases force the rings against the cylinders and the lower part of the piston groove and thereby seal the relatively large clearance between piston and liner. The top ring is subject to the highest pressure loading and thus suffers the most wear and has the largest contribution in friction.

The"rotating sleeve engine"is an invention that can significantly improve the lubrication conditions of the piston and piston rings, eliminate or significantly reduce wear and significantly reduce piston assembly friction.

In reciprocating piston engines, the piston linear speed is reduced to very low values at the regions with proximity to top and bottom dead centers. In those parts of the stroke, the sliding speed between the compression rings and the liner is insufficient for the maintenance of hydrodynamic lubrication. The protective lubricant film gradually breaks down and metal to metal contact occurs. The high cylinder pressure during the compression and power strokes loads the compression rings further, intensifying the phenomenon and expanding the portion of the stroke where the metal to metal contact occurs. Thus, localized wear on the liner around the dead centers and especially at the top is typical after prolonged engine operation. At the regions around the mid portion of the stroke, the piston speed reaches sufficient values for the hydrodynamic lubrication regime. The protective lubricant film prevents metal to metal contact, reduces the friction coefficient by up to two orders of magnitude, and essentially eliminates wear.

This can be verified by the fact that the mid portion of the liner is always free of wear.

Numerous frictional experiments reveal increased piston assembly friction around the dead centers due to the described phenomenon.

The above phenomenon is further illustrated by the Stribeck diagram shown in Figure 10 as presented by Irving J. Levinson, Machine Design. This diagram shows the friction coefficient between two sliding surfaces in the presence of lubricant as a function of the "duty parameter"which is defined as the product of sliding speed and lubricant viscosity divided by the normal contact pressure of the surfaces. When two surfaces slide in the presence of lubricant, three possible modes of lubrication are possible. At very low sliding speed and high normal load, boundary lubrication is present. Metal to metal contact is unavoidable. Due to surface adhesion, high level of friction and wear is present. As sliding speed and thus the duty parameter increases, hydrodynamic oil film pressure builds up, supporting a larger portion of the normal load. Thus, the two surfaces are gradually separated by the oil film with less and less asperity contact and reduced adhesive wear (mixed regime). Finally, at higher sliding speeds (duty parameter values of 50 or higher according to the Figure 10), the hydrodynamic pressure supports the entire load resulting in full separation. The metal to metal contact as well as wear are eliminated. In the part of the cycle when the piston approaches a dead center, the sliding speed approaches zero. Furthermore, when the piston is in proximity to the top dead center, compression-expansion stroke, the high cylinder gas pressure increases the normal load between the liner and the piston rings (which are practically pressure activated sealing devices) further reducing the value of the duty parameter. The result is that for a significant portion of the cycle, the duty parameter falls below the value of 50, with the corresponding high friction coefficient and level of wear.

The cylinder liner (also called the"sleeve") of the rotating sleeve engine rotates with the objective of maintaining a non zero sliding speed and large values of duty parameter throughout the stroke. According to the Stribeck diagram, the friction coefficient is reduced by almost two orders of magnitude for that particular portion of the stroke. The rotation can be achieved via gear mechanisms from the crankshaft (similarly to a distributor or injection pump). For best results, the magnitude of the rotation needs to be high enough in order maintain the hydrodynamic lubrication regime between the compression rings and liner, even when the piston linear speed is zero and the cylinder pressure is at its maximum value.

In conventional engines, the rings must be free to rotate to minimize localized ring wear.

However, both blowby and, for spark ignition engines, hydrocarbon emissions are affected by the relative azimuthal positions of the end gaps of the compression rings (Roberts and Matthews, 1996). When the rings are free to rotate, the engine designer cannot take advantage of these dependencies to help control blowby and hydrocarbon emissions. For the rotating sleeve engine, the rings can be pinned to prevent their rotation (which is no longer required to minimize wear).

In order to further investigate the feasibility of hydrodynamic lubrication just due to sleeve rotation, the Reynold's partial differential equation as shown by Hamrock (1994) was solved numerically in a situation that simulates a stationary piston ring subject to cylinder gas pressure while the liner rotates. The objective of the simulations is to explore the magnitude of the average hydrodynamic pressure obtainable by different liner sliding speeds and different ring profiles with a constant film thickness. The value of that pressure represents the maximum gas pressure that can be supported by the ring and still maintain the assumed film thickness. This pressure is the cylinder pressure at top dead center (TDC) compression stroke and is nearly equal to the peak cylinder pressure. The constant film thickness eliminates the contribution of squeeze film lubrication in the hydrodynamic film pressure and thus represents the worst case scenario for the rotating sleeve engine. It is as if the piston stays at top dead center indefinitely while the top compression ring is constant loaded with high gas pressure The value for the lubricant viscosity was for a 20W oil as given by Hamrock (1984). This is a low viscosity lubricant that minimizes the hydrodynamic losses at mid stroke. A flat piston ring profile was assumed with surface irregularities as the only means for pressure build- up. This phenomenon is called"microhydrodynamic lubrication"Hamrock (1984). The surface irregularities were set equal to the combined surface roughness used by the ring- pack modeling performed by Tian and coworkers (1996) of 0.3 microns. The irregularities were assumed to be on the liner surface only (while the ring surface was assumed to be perfectly flat) and their shape was a 2 dimensional sinusoidal wave. With a 3 m/s liner sliding speed and a mean film thickness of 1 and 0.8 microns (within the range of hydrodynamic lubrication for this size of the combined asperity size according to Tian and coworhers) the average lubricant pressure was 16.03 and 28.27 atm respectively. Furthermore, the average pressure demonstrated an almost proportional variation with liner speed.

In order for the lubricant film of 0.8 microns minimum film thickness and with low viscosity oil to be able to support the typical peak pressure for spark ignition engines of about 50 atm, liner speeds of over 6 mls are necessary. For heavy duty engines where the peak cylinder pressure can reach 100 atm, even higher speeds would be required.

Note that if the liner speed is not sufficient for the peak cylinder pressure, the film thickness will further drop with some metal to metal contact at dead center. However, the boundary lubrication will be still confined to smaller part of the cycle, where the piston speed is nearly zero, and thus the energy losses due to boundary lubrication are still minimized.

In a preferred embodiment, a new ring profile is incorporated in order to enhance hydrodynamic pressure with lower liner speeds. In a conventional engine, the top compression ring is equipped with a barrel shape with the intention of creating converging surfaces (Figure 6A, left figure) that enhance the build up of hydrodynamic lubricant pressure due to the up and down motion. Note that the depth of the barrel is ony 20 or 30 microns, and thus the profile looks perfectly flat to the naked eye. For the rotating sleeve engine, the depth of this shape can vary in the peripheral direction periodically as shown in Figure 6B (right figure) in order to create converging surfaces in the direction of sleeve rotation. Note that the shape is actually curved rather than angular as shown in Figure 6. At a point of maximum depth, the barrel shape remams like a typical compression ring. However, this depth drops linearly with peripheral displacement, until the barrel shape is reduced to a perfectly flat surface. Then, suddenly, the barrel shape is reintroduced and the same process repeats periodically. The result is that multiple"wedges"or converging surfaces are formed that enhance hydrodynamic pressure due to sleeve rotation. The length of these wedges along the periphery of the ring range between 1 and 1.5 ring widths.

The Reynold's equation was solved again for the new profile. With a film thickness of 0.8 microns and a sliding speed of 3 and 4 m/s, the average lubricant pressure was 88.10 and 110.82 atm respectively. Smaller film thickness and higher sliding speeds yield even higher pressure. As discussed above, if the slight piston motion present around dead center and squeeze film lubrication are included in the problem, similar hydrodynamic film pressure and/or higher minimum film thickness can be achieved for the short period of time that piston sliding speeds are low and cylinder gas pressure is high, with lower liner rotational speeds.

The above simulations indicate that hydrodynamic lubricant film pressure can be created at TDC or BDC just by the surface irregularities, even for a moderate liner rotation (3m/s corresponds to 409 rpm for a 5.5 inch bore engine) and with a low viscosity lubricant With the proposed ring profile, the lubricant pressure can be dramatically increased even with a relatively large film thickness reaching the magnitude of typical peak cylinder pressure for heavy duty diesel engines at full load. Note that the proposed ring profile increases the flat portion of the ring improving sealing and increasing the effectiveness of squeeze film lubrication at dead center. Part of the converging surface that enhances hydrodynamic lubrication due to up and down motion has been sacrificed. However, results from the models by Tian and coworkers (1996), Lawrence (1988) as well as several experimental studies indicate that there is more than sufficient film thickness at mid-stroke for hydrodynamic lubrication. Furthermore, since at least some liner rotation will be retained at mid-stroke, the converging surfaces at the peripheral direction will remain active and substitute for the overall reduction of the barrel shape.

Moving sleeves have been proposed for use in prior internal combustion engine patents.

However, in the prior art, the objective of sleeve motion is replacement of the conventional poppet valves with intake and exhaust slots that are exposed by the motion of the liner. For example Giorgio in U. S. patent 5,482,011, discloses an engine design where a liner rotates inside the engine block. The liner is tightly fitted on the block and is provided with a port which is aligned with similar ports on the block for the intake and exhaust processes. The rotational speed of the liner is restricted to one half the crank speed due to timing requirements. Akira in U. S. patent 5,191,863 describes an engine design with a ported rotating liner used for intake and exhaust processes. Again, the liner's motion is restricted by timing.

Some of the existing moving sleeve engines can show some improvements in the piston lubrication in respect to a stationary liner when sleeve rotation occurs at the dead centers, even though that was not the objective of the invention. An example is the series of sleeve valve engines developed by Ricardo before WWII (U. S. patent unknown) which shows great similarity to the engines described by Giorgio and Akira. However, in all these designs, the sleeve motion can not be optimized for friction optimization due to timing restrictions. Furthermore, even though Ricardo and coworkers (1968) reported some potential benefits in that particular design in terms of piston friction, the tight tolerances between the liner and block throughout the liner external area (necessary mainly due to the sealing requirements of the port openings provided on both the sleeves and stationary cylinders) introduced large loads on the driving mechanisms, minimizing the potential frictional benefits.

The ported engine designs shown by Richardo and coworkers as well as other US patents are using oil for the lubrication of the outside sleeve surface. Relatively large quantities of oil can be expected to enter the intake and exhaust flows through the ports. However, the oil consumption and resulting hydrocarbon emissions will not be compatible with modem US regulations. If different means of lubrication of the sleeves of the ported engines is attempted in order to avoid the excessive oil consumption (i. e. solid lubricant or dry lubrication), the advantages on wear and low friction coefficient inherent in fluid lubrication will not be present. The present invention incorporates conventional valvetrain, which prevents oil from entering the flow of intake or exhaust gasses.

Background on Moving Cylinder Sleeves Moving sleeves are not a new or untried engine feature. Ricardo and Hempson (1968) describes in detail the highly successful"sleeve valve"engines developed in the period between WWI and WWII and during WWII, mostly for aircraft applications (spark ignited engines). The objective of that design was the replacement of the conventional valve train and poppet valves. The motion of the sleeve would expose intake and exhaust slots at the right time in the cycle. The main avantage for aero engines was the reduction of the frontal area of the engine by the elimination of the rocker and/or overhead cam mechanisms. The shape and motion of the sleeves was designed in order to optimize the port exposure. A crank rotating at half the engine speed was connected by a ball joint to the sleeve, causing it to reciprocate and twist (at top dead center compression stroke, the sleeve's motion was momentarily purely rotational). Other advantages of these engines included central spark plug placement for denotation reduction and volumetric efficiency improvements. A second set of rings was necessary in order to seal the sleeve-cylinder head gap. This ring-pack was installed on the cylinder heads and was stationary.

During the early 20's two experimental single cylinder engines were constructed in Ricardo's laboratory in order to further explore the potential of the sleeve valve concept One was a conventional poppet valve engine with 4 valves, the other a sleeve valve engine. The two units were similar in every other respect. During the test, it was discovered that a"puzzling"feature of the sleeve valve engine was its lower frictional losses, in spite of the larger number of moving parts and large rubbing surfaces. Ricardo attributed this effect to the possibility of elimination of piston ring boundary lubrication due to the continuous ring-liner motion and continuous fluid lubrication of the rings.

That theory was conBnned by the observation that the sharply localized wear, always found in the liners of poppet-valve engines at the point where the top piston ring comes to rest at top dead center, was absent in the sleeve valve. Later tests on large number of sleeve-and poppet-valve engines of various shapes and sizes indicated that the overall mechanical losses of the sleeve-valve engine were usually less than those of the poppet- valve. According to Setright (1975), the Bristol"Centaums", a radial 18 cylinder air cooled high performance aero engine is holding the record for the longest operation between overhauls for a piston aero engine (3000 hours). Furthermore, the Nappier Sabre, a 24 cylinder liquid cooled high performance sleeve valve aero engine could maintain its combat rating almost indefinitely due to improved piston lubrication, while engines of the period with similar or lower Brake Mean Effective Pressure (BMEP) could maintain combat rating for only 5 minutes.

Evidence from the sleeve valve engines also indicates that the maintenance of hydrodynamic lubrication of the compression piston rings by the liner motion is feasible even with the relatively low rate of liner rotation that those engines had when the piston was at top dead center (TDC) compression-expansion stroke. Thus, wear protection of piston rings and liner is also feasible. Also, in the previous design, the primary function of the sleeve motion on the sleeve valve engines was not friction reduction. The clearance between sleeve and block was held low for gas sealing purposes around the intake and exhaust ports that were drilled on the sleeves and block In the current invention, the sleeve-block clearance can be chosen for friction optimization. Drawings of the sleeve valve engines reveal large rubbing surface between the block and sleeve.

The current invention reduces this rubbing area (by the introduction of the journal bearings) to the minimum necessary for fluid lubrication, and optimizes the sleeve motion for piston friction as well as sleeve friction. Thus, much higher frictional benefits are possible. Also, the reduction or elimination of metal to metal contact at TDC in those sleeve valve engines was possible due to the relatively low peak cylinder pressure typical of a spark ignition engine (compression ratio of about 6.5: 1). In order to achieve the same effect in a turbocharged heavy duty diesel engine with a compression ratio of 15: 1 or higher and peak cylinder pressure of about 100 atmospheres, higher rates of rotation and/or the proposed ring profile will be necessary SUMMARY OF INVENTION The invention is directed to an internal combustion engine having one rotating sleeve per cylinder that is supported by two journal bearings. The objective of the rotation is to maintain the sliding motion between the piston rings and liner in order to maintain the hydrodynamic lubrication regime throughout the stroke. The rings will be held stationary by the friction between the piston grooves and rings. However, if necessary, the rings can be pinned on the piston to prevent their rotation. Unlike the typical sleeve valve engine, the surface area where tight tolerances between the sleeve and the block exist are confined to the minimum necessary for reliable hydrodynamic support of the sleeve, minimizing the friction due to liner rotation. The intake of the fresh charge (or fresh air in the case of a Diesel) and exhaust of the combustion products are accomplished via conventional popper valve arrangement. A flange on the upper part of the liner transfers the thrust loads to the cylinder head and the block. Furthermore, the flange reinforces the upper part of the sleeve preventing or minimizing the sleeve expansion due to high cylinder pressure that occurs when the piston is in proximity to top dead center (TDC) which could otherwise take up the clearance of the upper journal bearing. This has been reported by Ricardo to be a typical phenomenon in sleeve valve engines without a flange, causing a serious penalty in friction and wear between the sleeve and the block. A set of pressure activated sealing devices are installed on the upper part of the sleeve to prevent high pressure combustion gases from entering the space between the liner and the block. The journal bearings that support the sleeve are supplied with pressurized oil from the oil pump lubricating the bearings. Lubricant leakage from these journal bearings fills up the space between the liner and the block and also lubricates the sealing devices on the upper part of the sleeve and the flange which acts as thrust bearing. The excess lubricant is removed by oil rettrn passages and by direct leakage back to the oil pan. The oil that surrounds the liner will remove excess heat zoom the engine, and may be used as the sole coolant fluid of the engine. In fact, the motion of the liner will enhance heat transfer from the liner to the oil.

The sleeves are driven by the crankshaft via gear mechanisms. Since the motion of the sleeve is not related to port operation unlike typical sleeve valve engines, the motion of the liners is independent of valve timing. Thus the final gear ratio between the crank and the sleeves can be optimized for friction optimization and wear minimization. For example, a turbocharged diesel engine with high peak cylinder pressure and low operating speeds will have such a gearing that the sleeve speed will be high for a given crankshaft speed in order to ensure full hydrodynamic lubrication of the piston rings at TDC compression stroke. On the other hand, an engine with high operating speeds and lower peak cylinder pressure like a spark ignition engine needs lower sleeve speed for a given crank speed to ensure hydrodynamic lubrication, and excessive sleeve speed would only result in excessive hydrodynamic losses. Furthermore, in applications where frictional losses are more important than wear of piston rings and liner, the sleeve speed can be less than the minimum necessary for full hydrodynamic lubrication at dead center compression stroke. In that case, some metal to metal contact may be likely at certain operating conditions (full load) but the overall losses due to sleeve rotation will be still kept low. Still, the metal to metal contact between the top compression ring and the liner will be confined on the small part of stroke when piston speed is very close to zero. The overall transmission ratio between the crank and sleeve can be adjusted by changing the gear ratio of the relevant gears.

The sleeve speed can be constant (for simplicity) or with alternating value (without reversal) for friction optimization. In the latter embodiment, the objective of alternating the value of the sleeve rotation is that at the mid portion of the stroke, the piston speed is sufficient for hydrodynamic lubrication regardless of liner rotation. Therefore, liner rotation at that part of the cycle does not offer any significant frictional benefits as at around the dead centers. Thus, minimizing the rate of rotation will reduce the additional friction caused by the liner rotation. However, even at mid stroke, a minimum rotational speed of the sleeve will have to be retained (even when the piston is at its peak speed) in order to retain the hydrodynamic oil film in the journal bearings with some small expense in friction. Then, as the pistons slow down, the driving mechanisms will gradually accelerate the sleeves. When the pistons come to a full stop, the sleeves will be at their peak speed. Then as the pistons speed up again, the sleeves could gradually slow down at their minimum speed when the pistons are at their maximum. An eccentric driving mechanism or a geartain based on the"Geneva Wheel"achieves this motion.

The kinetic energy stored in the sleeves at peak angular speed will return to the system during deceleration. In a multi-cylinder engine, the sleeves could be driven by independent mechanisms. However. gears installed on the outside of the sleeve surface (either the top or the bottom) could mesh with the sleeves of neighboring cylinders eliminating the need for individual driving mechanisms. The above driving method is particularly attractive if the constant sleeve speed embodiment is chosen, or if the alternating sleeve speed is to be used on a four cylinder engine (or in any in-line engine with a flat crankshaft design) where all pistons reach a dead center at the same time. In the latter case, the alternating magnitude rotation needs to be created by one mechanism and the motion will be duplicated from liner to liner.

Figure 11 shows the optimum liner speeds for different engine applications as derived from the numerical solution of the Reynold's equation. Again, the lubricant viscosity assumed is for an SAE 20 oil, a low viscosity lubricant that has the potential of minimizing the hydrodynamic losses at mid-stroke. The minimum film thickness is 0.8 microns which corresponds to a relatively small thickness but lies within the limits of hydrodynamic lubrication (based on the relevant literature). However, if the film thickness drops below that value, metal to metal contact is eminent. The average lubricant pressure shown in Figure 11 indicates the maximum cylinder pressure that can exist at TDC compression/expansion stroke, without any reduction in the film thickness.

Note that the pressure at TDC compression/expansion stroke is very close to the peak cylinder pressure that typically occurs between 10 and 20 crank angle degrees after TDC.

The two lines shown in Figure 11 show a typical (lower) and the revised compression ring profile (upper curve).

The gear selection for the sleeves is done as follows. The most frequent operating engine speed and load are selected. The corresponding peak cylinder pressure needs to be determined. Then, referring to Figure 11 depending on which compression ring profile is selected, the sleeve needs to rotate with an angular velocity such that the inner liner surface and peak pressure are at a point on the corresponding line or just below it.

The equation that relates the liner surface speed with the liner rotational speed is: V = fTR where V is velocity (m/s), R is bore radius (m) and H is sleeve angular velocity (rad/s). Then, simple arithmetic will determine the overall ratio of teeth of the driving gears. This applies for both constant and alternating speed. For the latter case, the maximum liner speed needs to be properly selected and is the one that is relevant to Figure 11. For example, a spark ignition engine with most frequent operating speed of 3000 rpm, a peak cylinder pressure of 35 atm, and a conventional ring profile, requires a sleeve angular velocity such that the inner liner surface speed is at 4m/s. If the speed is less than that, the minimum film thickness will drop below 0.8 microns at TDC, full load, and some metal to metal contact and wear could occur at TDC. For a turbocharged diesel engine with most frequent operating speed of 1200 rpm, a peak cylinder pressure of 95 atm, and a revised compression ring profile, about 3.5 m/s or less is required.

Note that the revised compression ring is recommended, since the liner rotation requirements is minimized with reduced sleeve friction. In that particular example, at engine speeds of higher than 1200 rpm, the sleeve speed will increase proportionally since it is driven by gear mechanisms by the crankshaft, ensuring fluid film lubrication of the ring. At lower speeds, if the peak cylinder pressure at full load remains approximately similar, some boundary friction is likely. Thus, if wear elimination is of paramount importance for the application, higher sleeve speed may have to be selected with an expense in high speed hydrodynamic losses. However, in typical heavy duty engines, the turbocharger speed drops at lower engine speeds due to reduced exhaust flowrate (lower turbocharger speed), resulting in lower peak cylinder pressure.

Also note that prior ported engines with a conventional ring profile and with a peak pressure/sleeve speed point that falls above the dotted line of Figure 11, the film thickness at TDC would drop below 0.8 microns with metal to metal contact, increased piston and sleeve friction and wear being very likely. However, if a lubricant with higher viscosity was used, the metal to metal contact could have been minimized or avoided, with an expense in the hydrodynamic friction of the piston at mid-stroke and in sleeve friction. In the case of the sleeve valve engines described by Ricardo, the typical compression ratio of those engines was around 6: 1 due to detonation limitations, resulting to far lower peak pressure as compare to a modem heavy duty diesel engine with compression ratio of around 14 to 17: 1. Also, the modem multi weight moror oils were not available at the time.

BRIEF DESCRIPTION OF DRAWINGS Figure 1 shows the general side view of the cylinder of the rotating sleeve engine. Note that clearances have been exaggerated for clarity.

Figure 2 shows a detail of the sleeve driving mechanism Figure 3 shows the means of alternating the magnitude of the sleeve speed.

Figure 4 shows the way adjacent sleeve gears engage each other.

Figure 5 shows the sleeve/cylinder head sealing mechanism Figure 6A shows a conventional ring profile. The drawing is exaggerated for clarity.

Note that profile is curved rather than angular as shown in the figure.

Figure 6B shows the new ring piston profile for the rotating sleeve engine. Again, the drawing is exaggerated for clarity and profile is curved rather than angular as shown in the figure.

Figure 7 shows the sleeve driving mechanism for a 4 cylinder rotating sleeve engine.

Figure 8 shows a top view and two cross sections of a section of the new piston ring profile.

Figure 9 shows sections the conical seal from three different views.

DESCRIPTION OF PREFERRED EMBODIMENTS Referring to Figure 1 (clearances have been exaggerated for clarity), there is shown one of the cylinders of a four stroke internal combustion engine with a rotating sleeve 1 that is supported by two journal bearings 2. The intake of the fresh charge (or fresh air in the case of a Diesel) and exhaust of the combustion products are accomplished via conventional poppet valve arrangement (not shown) which are located in the cylinder head 3. Thus, unlike typical sleeve valve engines, the motion of the liners is independent of valve timing and can be optimized for friction and wear minimization. A flange 4 on the upper part of the liner transfers the thrust loads to the cylinder head 3 and the block 5.

Furthermore, the flange reinforces the upper part of the sleeve preventing or minimizing the sleeve expansion due to high cylinder pressure that occur when the piston is in proximity to TDC. A piston 34 is fitted within the rotating sleeve and is connected via a connecting rod (not shown) to the crankshaft (not shown). The piston 34 is equipped with a compression ring 35 pinned on the piston to prevent its rotation, scraper ring 36, also pinned on the piston, and oil control ring 37. A set of pressure activated sealing devices 6 are installed on the upper part of the sleeve to prevent high pressure combustion gases from entering the space between the liner and the block. In Figure 1, one pressure activated seal 6 is shown. This is a compression ring with spring load against the outside of its groove machined into the head 3 and also spring load against the flange 4. The journal bearings 2 and 3 that support the sleeve are supplied with pressurized oil from the oil pump lubricating the bearings via oil passage 7 that is machined in the block. The clearance for these bearings is within the range for typical journal bearings of about. 002 to. 004 inches. The diameter to length ratio for the journal bearings is at typical values between 3 and 4. The upper bearing is located as high as possible, close to the flange. The location of the second bearing is not so strictly defined, but its centerline is at a level well below the middle of the piston stroke. Lubricant leakage from these journal bearings fills up the space between the liner and the block 8 and also lubricates the sealing device 6 and the flange 4 which acts as thrust bearing. The oil that surrounds the liner the space 8 will remove part of the excess heat from the engine. In fact, the motion of the liner will enhance heat transfer from the liner to the oil.

The flange 4 can be provided with pads as on hydrodynamic thrust bearings in order to promote hydrodynamic lubrication pressures. The excess lubricant is removed by oil return passages 9 and 10 and by direct leakage back to the oil pan. Referring to Figure 2, the sleeve 1 is equipped with a gear 11 that engages gear 12 which is connected to shaft 13. Shaft 13 is driven by gear 15 attached to the shaft 13. Gear 15 engages gear 14 which is part of the crankshaft 38 (shown only partially). As explained in previous sections, the final gear transmission ratio between the crank gear 14 and the sleeve 1 is very important for friction and wear consideration and can be optimized for different applications. This can be accomplished by changing the ratio of teeth gears 14,15,12 and 11.

Alternative Embodiment: Unpinned Compression Ring In the present alternative embodiment, the ring is not pinned in its piston groove.

However, when the high cylinder pressure take place during compression, the friction between the lower piston groove will prevail preventing the ring's rotation. The benefit of this is the rotation of the rings at other parts of the stroke will maintain the groove clean and free of combustion deposits.

Alternative Embodiment: Alternating Sleeve Speed The liner motion can be continuous (constant angular speed for a given crankshaft speed) for simplicity or with alternating value for this alternative embodiment. In the second case (alternative embodiment), a mechanism alternates the magnitude (but not the direction) of the liner angular velocity to a higher magnitude at the portion of the cycle when the piston is in proximity to a dead center. However, when the piston is around the mid portion of the stroke where hydrodynamic lubrication would exist anyway, the liner slows down for frictional savings. This is achieved preferably by using the Geneva wheel concept. The shaft 13 of Figure 2 is broken up in two parts as shown in Figure 3.

Wheel 16 is connected to the part of the shaft that is connected to the crankshaft, and rotates at constant speed for a given crankshaft speed. Also, wheel 16 is equipped with four pivots that engage the slots of wheel 17 that is connected to the part of the shaft that drives the gear 12 and thus the sleeve. Simple kinematics can show that the resulting speed of wheel 17 is alternating between a high value (pivot of wheel 16 is at its closest to the center of wheel 17) and a low value (pivot of wheel 16 is at its farthest location from the center of wheel 17). The number of teeth between gears 14 and 15 need to have a ratio of 1: 2 in order to have the speed alternation in phase with piston motion.

However, the ratio of teeth between gears 11 and 12 can be still altered arbitrarily in order to achieve any peak sleeve speed required at a given crankshaft speed. In the alternating sleeve speed embodiment, the sleeve thickness needs to be the minimum possible in order to reduce the inertia loads on the driving mechanism. If carbon steel is to be used for the sleeve material, for SI and diesel engines, this thickness should be around 1/16 and 3/16 of an inch respectively. Note that when the typical peak pressures for each type of engine is applied to a sleeve of infinite length with the corresponding thickness, the elastic deformation is of the same order as a typical journal bearing clearance. However, the flange at the top of the sleeve will minimize that deformation to a much lower number. When the area of the sleeve that is surrounded by the lower journal bearing is exposed to cylinder gas pressure, the pressure is dissipated due to gas expansion to very low value.

Alternative Embodiment: Drivins Mechanism for Multi-Cvlinder Benzines In an alternative embodiment of the invention applied in a multi cylinder engine, gears are installed on the outside of the sleeve surface as shown in Figures 4 and 7 and mesh with the sleeves of neighboring cylinders eliminating the need for individual driving mechanisms. Gear 18 (which is identical to gear 11 as shown in Figure 2) is bolted or press fitted on rotating sleeve 1, while a similar gear 19 is attached to the neighboring sleeve 20. The two gears are meshed through a slot machined on the block 5. Figure 7 shows this embodiment as applied to a 4 cylinder engine. Crankshaft 38, supported by main bearings (not shown), is equipped at the left end with gear 14 that transfers mechanical energy to the gear 11 of sleeve 1 of cylinder #1 (#1 refers to the unit to the far left of the Figure) via shaft 13. Gear 19 is attached to sleeve 20 of cylinder #2 and meshes with gear 18 (which is the same part as 11). The sleeve 20 of cylinder #2 similarly engages the gear of sleeve 41 of cylinder #3, etc. The above driving method is particularly attractive, if the constant sleeve speed embodiment is chosen, or if the alternating sleeve speed is to be used on a four cylinder engine (or in any in-line engine with a flat crankshaft design) where all pistons reach a dead center at the same time.

With this design, only one driving mechanism transfening mechanical energy from the crank to the sleeves needs to be installed, and the sleeve motion propagates to all cylinders. Figure 7 illustrates a typical 4 cylinder engine crankshaft of the"flat"design which means that all the connecting rod bearing centers are on one plane. As a result, all cylinders reach a dead center simultaneously. Pistons 39 and 40 of cylinders 1 and 4 respectively are shown at the bottom dead center (BDC) position while the pistons (not shown) of cylinder #2 and #3 are at TDC. Note that the instant shown in Figure 7 is the part in the engine cycle where the maximum sleeve speed is reached if the alternating speed embodiment is applied.

Alternative Embodiment: Conical Sealins Device In an alternative embodiment, a conical pressure activated sealing device has been conceived. Referring to Figure 5, a conical ring 21 fits on an also conical groove 22 machined on the cylinder head 3. Just like typical compression rings, that seal also has an open end and the diameter when uncompressed is a little larger than the groove it is supposed to fit in. Therefore, when forced into groove 22, the spring force developed forces it against the outside wall of groove 22. Due to the inclined surfaces, a spring load develops between the lower surface of ring 21 and the upper flat surface of the rotating sleeve. When large pressures due to combustion are developed within the cylinder, the ring 21 is further forced against the above surfaces intensifying the seaL The ring 21 could be held stationary (pinned) with respect to the cylinder head 3 and slide against the upper part of the rotating sleeve or be held stationary with respect to the sleeve and rotate with respect to the cylinder head. The presence of lubricant and continuous rotation are ideal for hydrodynamic lubrication. Again, part of the sliding surface can be provided with pads (just like a hydrodynamic thrust bearing and similar to the proposed compression ring design) to promote the formation of hydrodynamic lubrication. Figure 9 shows the design of the conical seal with pads at the location appropriate when the seal is pinned on the cylinder head and slides in respect to the sleeve's flange. 45 shows an overall three-dimensional sketch of a section of the seal with the pads 46 imbedded in order not to compromise the sealing. 47 is a side view with the pads 46 shown with hidden lines. 46 is a view of the lower surface with pads 46 clearly visible. The depth of these pads is exaggerated in Figure 9 for clarity. Note that the simulations discussed above also apply for the lubrication of the sliding surface of conical sealing ring of Figure 5. However, due to the ring's location (further from the axis of rotation), this ring will enjoy a slightly higher sliding speed.

Alternative Embodiment: Satina for Hish Peak Cylinder Pressure The following embodiment is particularly desirable for an engine that develops high combustion pressures (i. e. turbocharged diesel engine). In such an engine, the cylinder sealing is more critical. Therefore, referring to figure 5, an additional compression ring 23 is fitted on groove 24 machined on the flange of the rotating sleeve. Any combustion gasses that escape the main conical seal 21 are trapped by ring 23. Since ring 23 isolates the conical ring 21 from oil coming from the area 8 (area between sleeve and cylinder), an additional oil supply 25 is provided to lubricate the conical seal as well as the interface of the upper part of the flange and cylinder head which acts as thrust bearing.

A similar oil passage to 25 is provided on the opposite side and acts as oil retum.

Alternative Embodiment: Adiusted Comnression Ring Profile Hydrodynamic lubrication requires some form of converging surfaces, in order for lubricant pressures to form and support the normal load. When the piston is at top dead center and the only motion that occurs is due to liner rotation, there is no apparent converging surface (the surfaces are parallel). However, in reality the surface irregularities of the honed rotating liner provide minor converging (a phenomenon called micro-hydrodynamic lubrication). However, in the following alternative embodiment, the ring profile can be adjusted in order to create converging surfaces just due to ring rotation without relying on the surface irregularities and thus reducing the necessary rotation for fully hydrodynamic lubrication. Referring to Figure 6A, the conventional ring profile 49 is made out of the following portions. A flat portion 26 that acts as a sealing surface between the liner and the ring. Wedges 27 and 28 act as converging surfaces during the up and down motion respectively to promote formation of hydrodynamic pressure. Please note that the drawing has been exaggerated for clarity and wedge shape is so small that cannot be seen by the naked eye. Also, the shape is actually curved rather than angular as shown in Figure 6. The new ring profile 50 (Figure 6B) still maintains a flat portion 29 and is also provided with similar wedges at either edge 30 and 31. However, the depth of these wedges changes in the peripheral direction creating wedges in that direction as well. When the minimum wedge depth is reached, the same pattern is repeated. Figure 8 shows the top view 42 of a section of the new compression ring. Two cross sections, 43 and 44 are also shown in order to demonstrate the variation of the wedge depth. The result is that converging surfaces in the direction of sleeve rotation are effectively created. Thus, the pressure support of the ring is increased for a given liner rotational speed. This embodiment is particularly desirable for diesel engines where the peak cylinder pressure is very high. Simulations show that with the proposed ring profile, the lubricant pressure can be indeed dramatically increased even with a relatively large film thickness reaching the magnitude of typical peak cylinder pressure for heavy duty diesel engines at full load. Note that the proposed ring profile increases the flat portion of the ring improving sealing and increasing the effectiveness of squeeze film lubrication at dead center. Part of the converging surface that enhances hydrodynamic lubrication due to up and down motion has been sacrificed. However, results from the models by Tian and coworkers (1996), Lawrence (1988) as well as several experimental studies indicate that under most operating conditions, there is more than sufficient film thickness at mid-stroke for hydrodynamic lubrication. Furthermore, since at least some liner rotation will be retained at mid-stroke, the converging surfaces at the peripheral direction will remain active and substitute for the overall reduction of the barrel shape.

Alternative Embodiment: Adiusted Ring Profile Annlied to Conical Ring In another alternative embodiment, a profile using the above concept can is applied on the pressure activated sealing devices. In other words, ring 21 and 23 (Figure 5) have pads machined along their sealing surface in order to promote hydrodynamic lubricant pressure build up and avoid metal to metal contact. The pads are imbedded in the sliding surface of ring in order to allow tight sealing clearance in the rest of the surface. Note that the above simulations also apply for the lubrication of the lower flat sliding surface of conical sealing ring of previous embodiments. However, due to the ring's location (further from the axis of rotation), this ring will enjoy a slightly higher sliding speed. In addition to the surface irregularities, pads may be necessary to increase the film thickness when high gas pressure is encountered. Those pads can be similar in shape to typical hydrodynamic thrust bearings, but may have to be imbedded in the flat surface in order not to interfere with the sealing action. Also note that squeeze film lubrication will apply on this ring as well since the film thickness will be high before the gas pressure is raised.

Alternative Embodiment: Turbocharzed Diesel Benzine The following alternative embodiment describes the invention as applied to a turbocharged diesel engine with the maximum torque (and thus maximum peak cylinder pressure) at 1200 rpm. The typical maximum pressure at the peak torque speed at full load for such an engine is about 100 atm. Referring to Figure 11, the necessary liner speed in order to completely protect the liner and rings from metal to metal contact is around 3.5 m/s, if the proposed compression ring profile is to be used. For a 5.5 inch bore heavy duty engine, the resulting sleeve speed is 478 rpm. In the constant speed liner case, the gear ratio needs to be such that the gear 11 (Figure 7) attached on sleeve will spin at 478 rpm when crank gear 14 spins at 1200. Thus, the necessary condition for the number of teeth of gears 11,12,15, and 14 is: <BR> <BR> <BR> <BR> n" n, 2-478<BR> <BR> <BR> <BR> <BR> <BR> nt51200 Alternative Embodiment: Diesel Engine with Alternating Sleeve Speed As an alternative embodiment, the variable sleeve speed will be applied on the diesel engine of the previous embodiment. In this case, the maximum instantaneous sleeve speed occurring when the piston is at a dead center needs to be 478 rpm. Simple kinematics of wheels 16 and 17 will give the required speed of wheel 17 and 16 (Figure 3), and the corresponding gear ratios in order to achieve the liner speed of 478 rpm. The geometry of wheels 16 and 17 (distance of the two centerlines and radius of pivots on wheel 16) will be designed such that the required speed variation will be achieved. Note that the minimum speed should be chosen by two constraints. One is that the sleeve speed should not fall below a minimum necessary to retain the hydrodynamic film that support the sleeve. This minimum depends on the design of these bearings, viscosity lubricant used, the stroke to connecting rod length ratio and the cylinder pressure when the minimum speed occurs. The second is that an extreme speed variation could lead to excessive inertia loading on the driving mechanism which in turn will demand larger size gears to deal with the alternating inertia load.

Alternative Embodiment: Spark nition Engaine In the following alternative embodiment, the current invention is applied to a spark ignition engine, with a peak torque at 3000 rpm and a bore of 3.5 inches. The peak cylinder pressure at 3000 rpm is 50 atm. Referring to Figure 11, the required sleeve speed for full metal to metal protection at TDC is about 1.7 m/s. The corresponding sleeve speed is 365 rpm. For the constant sleeve speed, the gear ratio selection will be such that at 3000 crankshaft rpm, the sleeve gear spins at 365 rpm. The number of gear teeth selection is as above. For the alternating liner speed alternative embodiment, the maximum sleeve speed when the piston is at a dead center is 365 rpm.

Alternative Embodiment: Lubricant as Benzine Coolant In another alternative embodiment, the lubricant surrounding the rotating sleeve 1 in space 8 (Figure 1) can be used as the sole coolant of the engine. The flow rate needs to be sufficient to remove all excess heat from the cylinder walls. In conventional water cooled engines, water jackets are machined or cast in the cylinder heads in which water flows for cooling. In this embodiment, those passages are filled with lubricant. The oil pump powers the lubricant flow through these cylinder head passages removing excess heat. The whole oil flow rate is also pumped through an oil cooler of sufficient capacity in order to dissipate this waist heat into ambient air. Thus, the water pump, water radiator, and coolant hoses are replaced with a larger oil cooler and oil pump in order to compensate for the extra complexity of the system. The benefit of this embodiment is that the incremental cost of the rotating sleeve engine in respect to a conventional engine ir reduced.