Internal-combustion engines – Four-cycle – Engine cylinder having a reciprocating sleeve valve
Reexamination Certificate
1999-02-04
2001-09-18
Kamen, Noah P. (Department: 3747)
Internal-combustion engines
Four-cycle
Engine cylinder having a reciprocating sleeve valve
C123S059300, C123S0650VS, C123S08000R, C123S190120
Reexamination Certificate
active
06289872
ABSTRACT:
FIELD OF INVENTION
The invention pertains to an internal combustion engine being typically conventional but having a cylinder liner that rotates with the intention of reducing piston assembly friction and piston ring and liner wear.
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 efficiency 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
FIG. 1
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 Graph 1), 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 bellow the value of 50, with the corresponding high friction coefficient and level of wear.
Graph 1. Stribeck diagram
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 constantly 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
Huynh Hai
Kamen Noah P.
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