US20140360473A1 | 2014-12-11 | |||
US3765180A | 1973-10-16 | |||
US4237826A | 1980-12-09 | |||
US20090056331A1 | 2009-03-05 |
CLAIMS What is claimed is: 1. A method of operating an internal combustion engine having an intake and an exhaust manifold, and having a plurality of electronically controllable valves including an intake valve, an exhaust valve, and an electronically controllable fuel injector in each cylinder, each cylinder having a piston therein for oscillating through intake, compression, power and exhaust strokes comprising: a) providing an air rail, and providing an air valve in each cylinder; b) opening an intake valve coupling a first cylinder to the intake manifold during at least a part of an intake stroke of the respective piston; c) closing the intake valve opened in b) ; d) opening an air valve during the compression stroke of the first cylinder, and closing the air valve when the pressure in the first cylinder approaches or reaches the pressure in the air rail; e) after the end of the compression stroke and during the power stroke of the first cylinder and while fuel may be injected into the first cylinder without reaching a temperature above which NOx will be formed, but above the temperature to cause ignition of injected fuel, injecting fuel into the first cylinder; f) at the end of the power stroke, opening the air valve and then closing the air valve for the first cylinder during the exhaust stroke to exhaust part of exhaust gases from the first cylinder to the air rail, then opening the exhaust valve to couple the first cylinder to the exhaust manifold for the remainder of the exhaust stroke of the first cylinder . 2. The method of claim 1 wherein the injecting fuel into the first cylinder in e) comprises injecting fuel using a pulse injection. 3. The method of claim 1 further comprising after e) , at the end of the power stroke, leaving all valves closed and executing second compression and power strokes, and wherein in f ) , at the end of the power stroke is at the end of the second power stroke. 4. The method of claim 1 further comprising repeating b) through f) 5. The method of claim 1 further comprising: i) opening an intake valve coupling a second cylinder to the intake manifold during at least a part of an intake stroke of the respective piston, and then closing the intake valve; ii) opening an air valve during the compression stroke of the second cylinder, closing the air valve when the pressure in the second cylinder approaches or reaches the pressure in the air rail, and completing the compression stroke of the second cylinder; e) executing b) through f) . 6. The method of claim 5 wherein the second cylinder is the same cylinder as the first cylinder, and wherein i) and ii) are executed immediately before executing b) through f ) . 7. The method of claim 5 wherein the second cylinder is the same cylinder as the first cylinder, and wherein i) and ii) are executed multiple times, then immediately executing b) through f) . 8. The method of claim 5 wherein internal combustion engine is a multi-cylinder internal combustion engine and wherein the first and the second cylinders are different cylinders . 9. The method of claim 5 wherein internal combustion engine is a multi-cylinder internal combustion engine and wherein the first and the second cylinders are the same cylinder . 10. The method of claim 1 further comprising injecting a gaseous fuel into the mixture of air and exhaust gas in the air rail before or as the mixture of air and exhaust gas in the air rail flows into the first cylinder in d) . 11. The method of claim 10 wherein the gaseous fuel is or includes compressed natural gas. 12. The method of claim 10 wherein the gaseous fuel is or includes ammonia. 13. The method of claim 1 wherein the fuel is a diesel fuel . 14. The method of claim 1 wherein the fuel is ammonia. 15. The method of claim 14 wherein the ammonia is injected in liquid form. 16. A method of operating an internal combustion engine comprising : a) in first cylinder, at the end of a power stroke, delivering part of resulting exhaust gas in the first cylinder to an air rail and then exhausting the remaining exhaust gas; b) during an intake stroke of a second cylinder, intaking air; c) during the following compression stroke of the second cylinder, first delivering air and exhaust gas from the air rail to the second cylinder, then compressing the air and exhaust gas in the second cylinder during the rest of the compression stroke; d) after the end of the compression stroke of c) , injecting fuel into the second cylinder and executing a power stroke in the second cylinder; and e) repeating a) through d) at least one more time using the same or different cylinders. 17. The method of claim 16 wherein injecting fuel comprises injecting fuel using pulse injection. 18. The method of claim 16 wherein after d) , executing a second compression stroke and second power stroke in the second cylinder and then executing e) . 19. The method of claim 16 wherein the first and second cylinders are different cylinders. 20. The method of claim 16 wherein the first and second cylinders are the same cylinder. 21. The method of claim 16 further comprising: i) during an intake stroke of a cylinder, intaking air ; ii) during a compression stroke following i), delivering air to the air rail. 22. The method of claim 16 wherein i) and ii) are repeated on average more than once each time a) through d) are repeated. 23. The method of claim 16 further comprising injecting a gaseous fuel into the mixture of air and exhaust gas in the air rail before or as the mixture of air and exhaust gas in the air rail flows into the first cylinder in c) . 24. The method of claim 23 wherein the gaseous fuel is or includes compressed natural gas. 25. The method of claim 23 wherein the gaseous fuel is or includes ammonia. 26. The method of claim 16 wherein the fuel is a diesel fuel. 27. The method of claim 16 wherein the fuel is ammonia. 28. The method of claim 16 wherein the ammonia is injected in liquid form. 29. A method of operating an internal combustion engine comprising: a) in first cylinder, at the end of a power stroke, delivering part of resulting exhaust gas in the first cylinder to an air rail and then exhausting the remaining exhaust gas; b) during an intake stroke of a second cylinder, intaking air; c) during the following compression stroke of the second cylinder, first delivering air and exhaust gas from the air rail while or after delivering a gaseous fuel to the air rail, to the second cylinder, then compressing the gaseous fuel, air and exhaust gas in the second cylinder during the rest of the compression stroke to obtain compression ignition at or near the end of the compression stroke; d) after the end of the compression stroke of c) , executing a power stroke in the second cylinder; and e) repeating a) through d) at least one more time. 30. The method of claim 29 wherein after d) , executing a second compression stroke and second power stroke in the second cylinder and then executing e) . 31. The method of claim 29 wherein the first and second cylinders are different cylinders. 32. The method of claim 29 wherein the first and second cylinders are the same cylinder. 33. The method of claim 29 further comprising: i) during an intake stroke of a cylinder, intaking air; ii) during a compression stroke following i), delivering compressed air to the air rail. 34. The method of claim 29 wherein i) and ii) are repeated on average more than once each time a) through d) are repeated. 35. The method of claim 29 wherein the gaseous fuel is or includes compressed natural gas. 36. The method of claim 29 wherein the gaseous fuel is or includes ammonia. |
BACKGROUND OF THE INVENTION
1. Field of the Invention
The present invention relates to the field of
compression ignition engines.
2. Prior Art
Internal combustion engines currently in use have a single compression with a fixed mechanical compression ratio. In the case of a spark ignition engine, the fuel/air ratio must be maintained within a narrow range at idle, for low power settings and for high power settings. This is achieved by throttling the air flow into the engine, which indirectly limits the effective compression ratio to much lower values under such operating conditions. Such limiting of the effective compression ratio by throttling is not generally desired, though is a practical and effective way of limiting the output of the spark ignition engine under idle and low power settings. In the case of compression ignition engines, the fuel/air ratio used is not limited by such
considerations, and accordingly, may be varied from values suitable for idle conditions up to maximum power settings, so that in general the compression achieved is that defined by the mechanical ratio, though with any supercharger on the engine effectively providing a somewhat higher compression than the mechanical ratio itself. However, since most superchargers are turbochargers powered by the exhaust gas from the engine, their effect will be minimal except for higher power settings. Since their operation is dependent on the power settings, and turbochargers do not respond
immediately to changes in power settings, they are not effective as a primary control of effective compression ratio in the engine .
Further, in present compression ignition engines, because of the single compression together with the high compression ratio required to attain self-ignition of even diesel fuel (including biodiesel fuel) before the air in the combustion chamber cools, ignition is initiated at or just past the piston top dead center position, which results in the maximum combustion chamber pressure occurring at a crankshaft angle that is not favorable for the conversion of the combustion chamber pressure to mechanical crankshaft output. Pulsing of the fuel injection helps extend the combustion process over a greater crankshaft angle, though still normally the initial pulses occur at an undesirably small crankshaft angle. Also the temperature achieved even at the top dead center position of a piston is not adequate for self-ignition of other, otherwise attractive fuels to use, such as compressed natural gas (CNG) and ammonia (NH3) .
Thus, it would be desirable to have an engine with a directly and controllable effective compression ratio, not only for engine operation optimization, but also to allow the use of various fuels, which fuels may have substantially differing ignition characteristics. Further, it would be desirable to have an engine in which the combustion can be initiated over a relatively wide variation in crankshaft angle and extended through a desired and controllable
crankshaft angle range, again for engine operation
optimization . BRIEF DESCRIPTION OF THE DRAWINGS
Fig. 1 is a schematic drawing of a piston type engine having intake and exhaust manifolds, as well as an air rail for temporary storage of pressurized air at an elevated temperature, and provisions for battery energy storage and hydraulic energy storage.
Fig. 2 is a schematic diagram illustrating four-stroke operation for an exemplary method of operating an engine like that of Fig. 1. Fig. 3 illustrates a staggered operation of each of the four cylinders of Fig. 1, though with the cylinders labeled in accordance with their sequence of operation and not necessarily in accordance with their physical position within the exemplary engine of Fig. 1. Fig. 4 illustrates an exemplary operation of an eight cylinder engine in an eight-stroke cycle.
Fig. 5 illustrates an exemplary operation of a six cylinder engine in a six-stroke cycle.
Fig. 6 illustrates an exemplary operation of an eight cylinder engine in an eight-stroke cycle that includes a two- stroke reburn.
Fig. 7 illustrates an exemplary control system for an engine in accordance with the present invention.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
First referring to Fig. 1, an embodiment of the present invention may be seen. In particular, in that Figure, a 4-cylinder internal combustion engine 20 incorporating the present invention is schematically illustrated. As shown in that Figure, each cylinder has a pair of intake valves I 22, an exhaust valve E 24, and what will be referred to herein as an air valve A 26. In addition, each cylinder has a fuel injector F 28, typically for injecting a liquid fuel such as a diesel fuel at the appropriate time, as shall be described in greater detail later herein.
The intake valves I 22 of each cylinder are coupled to an intake manifold 30 and the exhaust valves E 24 are coupled to an exhaust manifold 32. These manifolds may be of conventional design, and in fact, if the invention is being applied to an existing engine or at least an existing engine design, may be in accordance with that engine or design generally, possibly with minor modification of the exhaust manifold or engine porting to use a single valve for the exhaust valve E 24 instead of two valves, E 24 and A 26 for exhaust valves as that engine design may have intended to be used. Also shown in Fig. 1 is an air rail 34, linking the air valves 26 of all cylinders. Also schematically shown in Fig. 1 is the engine crankshaft 36, an electric
generator/motor 38 coupled to the crankshaft, and a battery 40 which may be coupled to a solar panel 42 as well as a hydraulic pump/motor 44, a hydraulic accumulator 46 and a hydraulic reservoir 48. In an actual engine design, the air rail 34 physically may be a separate plate-like manifold, both the air rail 34 and the exhaust manifold 32 being bolted to the engine block in the normal manner for an exhaust manifold, simply using longer bolts. In the engine of Fig. 1, the valves 22, 24 and 26 are electronically controlled, hydraulically actuated valves generally, but not necessarily limited to those of any of U.S. Patent Nos. 5,638,781, 5,713,316, 5,960,753, 5,970,956, 6, 148, 778, 6, 173, 685, 6, 308, 690, 6, 360, 728, 6, 415, 749,
6, 557, 506, 6, 575, 126, 6, 739, 293, 7, 025, 326, 7, 032, 574,
7,182,068, 7,341,028, 7,387,095, 7,568,633 7,730,858,
8,342,153 and 8,629,745, and U.S. Patent Application
Publication No. 2007/0113906. Similarly, the fuel injectors 28 are electronically controllable, hydraulically actuated fuel injectors which may be in accordance with any of U.S. Patent Nos. 5,460,329, 5,720,261, 5,829,396, 5,954,030, 6, 012, 644, 6, 085, 991, 6, 161, 770, 6, 257, 499, 7, 032, 574,
7,108,200, 7,182,068, 7,412,969, 7,568,632, 7,568,633,
7,694,891, 7,717,359, 8,196,844, 8,282,020, 8,342,153,
8,366,018, 8,579,207, 8,628,031, 8,733,671 and 9,181,890, U.S. Patent Application Publication Nos. 2002/0017573,
2006/0192028, 2007/0007362 and 2010/0012745, and
International Publication No. WO 2016/196839. Such
electronic operation provides full control of the timing and duration of operation of the valves as well as the fuel injector, and accordingly, provides the ability not only for a very flexible control of the engine, but further the capability of making minor corrections cycle to cycle during operation of the engine to maximize its performance,
typically for efficiency purposes at the then existing power setting. Other forms of electronically controllable engine valves and fuel injectors are also known, and may be used in embodiments of the present invention as desired.
Now referring to Fig. 2, a schematic illustration of the preferred operating cycle of the present invention may be seen. This operating cycle is a four stroke cycle comprising an intake stroke I, a compression stroke C, a power stroke P and an exhaust stroke E. Also illustrated in Fig. 2 is the operation of the intake valves 22, labelled I in the Figure, the operation of the exhaust valves 24, labelled E in that Figure, and the operation of the air valve 26 of Fig. 1, labelled A in Fig. 2. Any of these valve designations may be followed by an 0 for opening or open and a C for closing or closed . Assume for the moment that the engine has been running for a while and is up to operating temperature. Thus, at the end of the power stroke P when the piston is at the bottom dead center position B2, the respective cylinder will contain hot exhaust gas still at a substantially elevated pressure. When this occurs, the respective air valve 26, A in Fig. 2, is opened for part of the exhaust stroke and then closed to direct part of the exhaust gas to the air rail A, raising the temperature and pressure of the air rail A, after which the exhaust valve E itself is opened and then closed at the end of exhaust stroke Tl, thus capturing part of the hot exhaust gases in the air rail 34 and then completing the exhausting of the respective cylinder. Then when the exhaust valve E closes, a conventional intake stroke I is executed, with the respective intake valves I 22 being opened at or near the top dead center position Tl and then closed at or near the bottom dead center piston position Bl . After this, during the early part of the compression stroke C, the air valve A 26 for the respective cylinder is opened and then somewhat later in the compression stroke is closed, with compression proceeding to the top dead center position T2 between the compression C and power stroke P. Typically, because of the inclusion of the hot exhaust gases during the exhaust stroke E, the
temperature in the combustion chamber will be too high for injection of fuel at the top dead center position T2 without creating NO x , though as the pressure and temperature in the combustion chamber decreases during the power stroke, the respective fuel injector F 28 may be operated to inject the fuel F, preferably using a series of pulse injections, before the temperature in the combustion chamber falls below the compression ignition temperature of the injected fuel, with the pulsed injections starting well after the respective piston has passed its top dead center position and the crankshaft is at a more favorable crankshaft angle for mechanical power production, and sustaining combustion over a much wider crankshaft angle than normal, for the most
efficient conversion of the pressure energy in the combustion chamber to mechanical energy. Having now given an overview of the operating cycle of
Fig. 2, certain additional characteristics thereof will now be set forth. In particular, as stated before, during an intake stroke, the air taken in from the intake manifold 30 through intake valves 22 will have a volume equal to the full displacement of the respective cylinder. Air is delivered in part to the air rail 34 and in part to the combustion chamber of the respective cylinder or another specific cylinder, through the operation of the respective air valve 26.
Consequently, since the volume of air taken in during the intake stroke is equal to a full displacement of the
respective cylinder plus additional exhaust gas that was added to the air rail 34 during the exhaust stroke as
hereinbefore described, the total amount of air (and exhaust gas) delivered to the combustion chamber during the
compression stroke must be greater than the amount of the intake air during the intake stroke I . Accordingly, the pressure in that cylinder during the compression stroke C will be higher than it would have been had some of the exhaust gas not been passed to the air rail 34 during the exhaust stroke E. Further of course, that mixture being compressed during a compression stroke C will also be
considerably hotter than it would have been, but for the part of the exhaust gas being passed to the air rail 34 during the exhaust stroke, allowing the injection of the fuel through the fuel injectors 28 well after the respective piston has passed the top dead center position, allowing the efficient conversion of the pressure energy to mechanical energy through the pistons' coupling to the crankshaft as
hereinbefore described.
In the engine just described, the amount of exhaust gas added to the air rail 24 in essence determines the effective compression ratio of the engine, which because of the heating provided by the exhaust gas in the air rail and in the air and exhaust gas mixture injected into the combustion cylinder early in the compression stroke, results in both a
controllable compression ratio and a generally high effective geometric compression ratio because of the temperature of the gas mixture during compression leading up to the power stroke. Of course the effective compression ratio and the amount of exhaust gas recirculation is fully controllable because of the electronic control provided for the engine valves . Now referring to Fig. 3, the operating cycle illustrated in Fig. 2 is illustrated for all four cylinders of the engine of Fig. 1, with the second cylinder having the same four stroke cycle as the first cylinder, though delayed by a stroke, the third cylinder delayed by another stroke, and finally the fourth cylinder delayed still a further stroke, so that the cycle may repeat in a still further four stroke cycle for cylinder one. In that regard, while the cylinders are labeled 1, 2, 3 and 4, that may or may not be in
accordance with the physical layout of the cylinders, but rather may be indicative of the succession of the power stroke. Typically the operation of the cylinders is staggered, so as to result in better dynamic balance of the engine .
Now referring to Fig. 4, the operating cycle for an 8-cylinder eight stroke engine may be seen. While these curves generally reflect combustion chamber pressures, combustion chamber temperatures generally reflect a similar profile. In essence, this operating cycle provides two compression strokes C2 and C3 for delivering air to the air rail, with the final compression stroke CI receiving a mixture of air and exhaust gas early in that compression stroke so that a still much higher compression ratio may be achieved. Note that using the illustration in Fig. 3 as an example, successive cylinders (logically, but not necessarily physically) would be shifted one stroke so that Fig. 4 would repeat at the end of the eight strokes.
Similarly, in Fig. 5 a six stroke cycle for a 6-cylinder engine is illustrated. In this case only two compression strokes are executed overall, not three as in the 8-cylinder eight stroke engine, the second compression stroke CI being the compression stroke leading to the power stroke and pulsed injection of the fuel after the respective piston reaches top dead center and retreats therefrom, again providing a more efficient crankshaft angle for conversion of the pressure energy to mechanical energy. However, it should be noted that because of the flexibility in the valve timing and operation, one might run the six stroke cycle in the
8-cylinder engine, or vice versa, as well as possibly running the four stroke cycle of Fig. 2 in the 6-cylinder or
8-cylinder engines. In that regard, it is the total
flexibility in valve timing and duration, as well as the total flexibility in the fuel injection, that allows the operation of various cycles from time to time without
restriction on the engines of the present invention.
Now referring to Fig. 6, a further embodiment of the present invention method may be seen. This embodiment illustrates an exemplary operation of an eight cylinder engine in an 8-stroke cycle that includes a 2-stroke reburn. As may be seen therein, the power stroke P is followed by a pair of reburn strokes R, which, with all valves closed, simply recompresses the respective cylinder contents and then allows the same to expand, which reheats the contents of the cylinder to above the compression-ignition temperature of the fuel used and then allows the same to expand again as a second power stroke. The net effect of this is that any unburned fuel that there may be in the cylinder contents at the end of the first power stroke winds up being better mixed with the air remaining in the cylinder and of course heated substantially and vaporizing throughout the recompression (the first stroke of the reburn stroke pair) , so as to be ignited when the compression ignition temperature is reached. This assures that any fuel left unburned after the first power stroke because of the prior uneven mixing of the fuel and air is ultimately totally consumed. In that regard, one of the problems encountered with direct injection compression ignition engines is the fact that when the piston is at the top dead center position, or at least still close to that position to achieve the compression ignition, any fuel injection needs to be directed primarily outward from the injector so as to avoid spraying the piston top with fuel that will not then burn. Because the nozzles on fuel
injectors are fixed nozzles, as the piston moves downward, the further injection of fuel is too radial, so that the region close to the fuel injector may have fuel rich regions, whereas the regions closer to the top of the piston may be fuel lean regions. This of course would be particularly true for higher loads on the engine, but could occur under any operating conditions. Thus at the end of the two strokes of the reburn, provided the overall combustion chamber contents were not fuel rich, the excess air in the combustion chamber should adequately burn any remaining unburned fuel.
Note also that in Fig. 6, at the end of the expansion stroke of the 2-stroke reburn R, the cycle essentially picks up with the stroke labelled E in Fig. 4 for the original 8- stroke eight cylinder operation. Thus the 8-stroke operation of Fig. 6 with the two reburn strokes R is the same as that shown in Fig. 4, but with the two reburn strokes R occurring between strokes P and E, with the rest of the sequence shown in Fig. 4 applying, with the exception that one of the intake and compression cycles 13 and C3 is eliminated. In general, if the reburn stroke pair is used, the temperature achieved on the reburn compression stroke needs to clearly exceed the compression ignition temperature for the fuel used, though that is not normally a problem, and of course in all
instances the system must be controlled to avoid temperatures exceeding the temperature at which meaningful amounts of NO x will be generated, normally accepted as being generated at temperatures exceeding 2200°C. Accordingly, whether running on a liquid fuel such as a diesel fuel or ammonia, or running on a gaseous fuel such as compressed natural gas or ammonia, or running on a combination of a liquid and a gaseous fuel, this upper temperature limit must be recognized and
controlled .
Now referring to Fig. 7, a block diagram of an exemplary control system for engines in accordance with the present invention may be seen. As shown therein, a controller, generally a microprocessor based controller, receives various inputs, including engine operating conditions which may include engine RPM, engine temperature, air rail temperature and pressure, as well as environmental conditions such as atmospheric pressure and atmospheric temperature, and in some embodiments, atmospheric humidity. The controller itself would include random access memory and read only memory
(which could be or include electronically reprogrammable memory) for storing instructions in a non-transitory machine readable medium which, when executed by the controller, would cause the various methods of the present invention to be executed. Also coupled to the controller would be a series of lookup tables (LUTs), which lookup tables would provide sequences of engine operation based on a power setting and the other parameters provided to the controller, to generally set the best operating sequences (such as those of Figs. 3, 4 and 5, or others one might choose to use) to nominally achieve the best engine performance, normally the best operating efficiency, unless there is some other overriding concern, such as by way of example, a present need for maximum power output. In addition, the software may include sequences for iterative adjustment of the various engine operating parameters to further maximize performance, typically in an iterative process using feedback of such parameters as engine power setting, fuel flow rate to the engine and output power of the engine.
The output of the controller of course is used to control each individual injector as well as the various engine valves hereinbefore described. With the electronic control of the fuel injector and engine valves, the timing and duration of the operation of each are totally
controllable so that the controller typically has total control of the engine responsive to the power setting. In that regard, that power setting may be by way of an input from a vehicle operator to duplicate a vehicle accelerator setting, to a generator type engine application wherein the power setting is automatically controlled to maintain the average AC power output frequency at the desired level in spite of momentary disturbances therein that result from changing generator loads, etc.
It was previously pointed out that the amount of exhaust gas added to the air rail 24 in essence determines the effective compression ratio of the engine. In addition, if desired, the compression ratio could be varied by also controlling the amount of air introduced into the engine during the initial intake stroke by opening the intake valves at the beginning of the respective intake stroke, but closing the intake valves before the end of the intake stroke. Thus as may be seen, there are various ways that the various parameters that may be controlled by the controller to achieve various desired effects. Similar but different effects could be achieved by closing the air valve A earlier (or later) than as described herein. Having now generally described an engine in accordance with the present invention and various aspects of its
control, various other aspects and highlights thereof will now be described. First, of course, is the fact that because one can achieve substantially any compression temperature prior to ignition of a fuel, that capability introduces the possibility of using gaseous fuels such as CNG or even NH 3 in gaseous form. Such fuels, while being introduced through the gaseous fuel injectors GF prior to the final compression, would self-ignite when the required temperature was reached, so their ignition at the proper time would be dependent upon control of the various parameters of the engine as
hereinbefore described, and could not really be delayed past approximately the top dead center position of the respective piston. Consequently it is desirable to sense ignition in at least one cylinder, such as by way of a pressure transducer, subsequently described, or by way of a microphone or
accelerometer to sense the engine response to ignition.
Of course ammonia can also be introduced as a liquid through a liquid fuel injector F, which then would not be subject to the same restrictions as introducing the same as a gaseous fuel. In particular, the injections of ammonia in liquid form could be delayed past the top dead center
position of the respective piston, provided the self-ignition temperature for the N¾ was sustained until injection, and of course pulse injection could be used with such a liquid fuel. The pressure of the ammonia may have to be substantially elevated above its normal storage pressure by a high pressure pump to maintain the ammonia as a liquid at the temperatures it may be subjected to, though that energy should be
retrieved through the phase change that will occur
substantially immediately upon injection into the combustion chamber. Also hydrocarbon fuels other than diesel fuel may be injected in liquid of gaseous form, such as even gasoline.
Further, the electric generator/motor 38 and battery 40, as well as the hydraulic pump motor 44 and accumulator 46 with its reservoir 48, have been previously mentioned. The size of the electric generator/battery 40 may range from something adequate to power accessories and a starter motor to a substantial battery bank. Similarly, the hydraulic pump/motor 44 and the accumulator 46 may be sized to power some accessories and the electronically controlled hydraulic valve actuation system and the hydraulically powered but electronically controlled fuel injectors, to a hydraulic/pump motor 44 and accumulator 46 of substantial size so that either the battery or the hydraulic accumulator, or both, may store substantial energy when the engine is used for braking, for bursts of overall power output when needed, or even to allow operation of the engine at a high power, high
efficiency operating point, even when that full power is not needed, but instead to save the excess engine output for later operation of the vehicle or whatever other system is being powered by the engine on battery power or accumulator power, or both, as the case may be. The electrical storage, of course, can be enhanced by the solar panel 42.
Because of the complete electronic control of the engine, the number of strokes for a full operating cycle can be the same as the number of cylinders in the engine as shown in Figs. 3, 4 and 5, though that is not a limitation of the invention. In particular, the number of strokes per complete engine operating cycle may be different from the number of pistons in the engine, and may range from less through more strokes per complete engine operating cycle than the number of cylinders in the engine. Further, the number of strokes per complete engine cycle need not be constant, but rather may vary engine operating cycle to engine operating cycle. By way of example, one might have an eight cylinder engine that operates in a four-stroke engine operating cycle in accordance with Fig. 3, then operating in a six-stroke engine cycle in accordance with Fig. 5, perhaps oscillating between those two operating cycles, or even operating in accordance with a four-stroke engine operating cycle for two engine operating cycles and then going to a single six-stroke engine operating cycle, and then back to a four-stroke engine operating cycle. When so doing, it is preferable to attempt to maintain the same firing order and to maintain the engine as well balanced as possible throughout such operation, though even this is not a real limitation of the invention. The net effect of this is to allow not only one or more additional compression strokes like those of 12 and C2, and 13 and C3 per compete engine cycle, but also on average, allow one or more partial additional compression strokes. Note also that the air rail may include a substantial
compressed air energy storage capacity, if desired.
Finally, while the invention has been disclosed and described herein with respect to an exemplary embodiment comprising a crankshaft type piston engine, a free piston engine assembly, such as in accordance with U.S. Patent Nos. 8,596,230, 9,206,738 and 9,464,569, may be used if desired. The use of a free piston engine has a number of advantages, including the fact that a free piston engine of N cylinders may be operated on all N cylinders or any lesser number, as may be most efficient at the time. Similarly, the idle condition for a free piston engine may in fact be an engine off condition, as a free piston engine may essentially start from an off condition to a full operating speed in a single engine operating cycle. Accordingly, in essence a free piston engine introduces still further degrees of flexibility in the engine operation.
For starting purposes, one might start the engine using a conventional four stroke diesel cycle with no exhaust gas mixing, or alternatively, using a six stroke or eight stroke cycle to at least compress the intake air to a substantially higher pressure, and as is most important for starting a cold engine, to a substantially higher temperature in the air rail and thus a substantially higher temperature at the top dead center position preceding the power stroke with the fuel being injected earlier, at or just after that top dead center position occurs, to be sure that ignition is achieved.
Because of that flexibility, the engine should more readily start in cold environments than a conventional diesel engine might start.
Thus, what has been disclosed herein is essentially a variable compression ratio engine that may recover some of the energy in the waste heat of the exhaust of a conventional engine and which converts the combustion chamber pressure energy to mechanical energy at the crankshaft output at more favorable crankshaft angles for higher efficiency energy conversion. Of course as mentioned before, minor iterative adjustments in the electronic controls for an engine in accordance with this disclosure allows the engine to
automatically seek its best operating characteristics during normal operation of the engine.
As previously mentioned, a pressure transducer can be used to detect the compression pressure at the top dead center position of a piston between the compression and power strokes and as an indicator of ignition. A pressure
transducer for each cylinder during testing of an engine is desirable, though one per engine for production purposes should be sufficient as a cost effective compromise. Of course engines of the present invention may also operate on various other cycles from time to time, which cycles may be unrelated to the present invention, or may include the present invention plus more. Conversely, additional engine instrumentation and/or control may prove useful, such as, by way of example special engine temperature sensors and control loops that will result in the controller using lower heat generating algorithms for operating the engine if overheating is threatening, or to put time limits on certain engine operation for similar reasons.
Thus the present invention has a number of aspects, which aspects may be practiced alone or in various combinations or sub-combinations, as desired. While certain preferred embodiments of the present invention have been disclosed and described herein for purposes of illustration and not for purposes of limitation, it will be understood by those skilled in the art that various changes in form and detail may be made therein without departing from the spirit and scope of the invention as defined by the full breadth of the following claims.
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