CA2620602C - Homogeneous charge compression ignition (hcci) vane-piston rotary engine - Google Patents
Homogeneous charge compression ignition (hcci) vane-piston rotary engine Download PDFInfo
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- CA2620602C CA2620602C CA2620602A CA2620602A CA2620602C CA 2620602 C CA2620602 C CA 2620602C CA 2620602 A CA2620602 A CA 2620602A CA 2620602 A CA2620602 A CA 2620602A CA 2620602 C CA2620602 C CA 2620602C
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Classifications
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
- F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
- F02B1/00—Engines characterised by fuel-air mixture compression
- F02B1/12—Engines characterised by fuel-air mixture compression with compression ignition
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
- F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
- F01C1/00—Rotary-piston machines or engines
- F01C1/30—Rotary-piston machines or engines having the characteristics covered by two or more groups F01C1/02, F01C1/08, F01C1/22, F01C1/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members
- F01C1/34—Rotary-piston machines or engines having the characteristics covered by two or more groups F01C1/02, F01C1/08, F01C1/22, F01C1/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F01C1/08 or F01C1/22 and relative reciprocation between the co-operating members
- F01C1/344—Rotary-piston machines or engines having the characteristics covered by two or more groups F01C1/02, F01C1/08, F01C1/22, F01C1/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F01C1/08 or F01C1/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the inner member
- F01C1/3446—Rotary-piston machines or engines having the characteristics covered by two or more groups F01C1/02, F01C1/08, F01C1/22, F01C1/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F01C1/08 or F01C1/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the inner member the inner and outer member being in contact along more than one line or surface
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
- F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
- F02B53/00—Internal-combustion aspects of rotary-piston or oscillating-piston engines
- F02B53/02—Methods of operating
-
- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
- Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
- Y02T10/00—Road transport of goods or passengers
- Y02T10/10—Internal combustion engine [ICE] based vehicles
- Y02T10/12—Improving ICE efficiencies
Abstract
For over a century, when three liters of gas are burned in a passenger car with a gasoline Spark-Ignition (SI), or Diesel Compression-Ignition (CI) piston engine, the 1st liter is wasted in heat, the 2nd liter is lost in throttling at intake manifold and/or gas kinetic energy at tail pipe (a loss that turbochargers and today's Atkinson thermodynamic cycle engine design with slightly increased expansion ratio (ER) tend to reduce), only the 3rd liter is used to propel the vehicle. This explains why, despite on-going technological efforts and billions of R&D dollars spent, thermal efficiency of most SI and CI piston-crankshaft engines stagnates at around 33%. The last three decade's rising gas prices and air emission standards have invited the automobile industry to study the more promising HCCI combustion mode.
Controlling such a fast combustion in conventional piston-crankshaft engines throughout a useful load-speed range has proven difficult so far, as combustion chambers are sized for deflagration on SI, or flame diffusion on CI or Diesel, but not for detonation. In four-stroke engines two crankshaft revolutions are required for each power stroke, and the shallow sinusoidal compression ratio (CR) variation inherent to the piston-crank kinematics near top dead center (TDC) tends to cause either misfiring when engine is cold or runs too lean (CR value insufficient for detonation near TDC) or severe knocking, excessive vehicle noise-vibration-harshness (NVH) and structurally intolerable pressure rise rate during transients, when engine gets hot or runs higher load at lower revolutions per minute (RPM) (premature detonation problem). As a result, most HCCI
concepts in the works are based on sophisticated sensor-computer-actuator piston-crankshaft engines that run on HCCI mode at part load or idle, and revert to the century-old thermally and environmentally less efficient, but structurally gentler, SI or CI mode at full load, or on cold start.
This invention aims to control HCCI under all operating conditions, thus to minimize engine air emissions and heat losses by means of a vane-piston engine: A radial vane pump with vanes positively guided by inner and outer cams, compresses gradually air, or a homogeneous air-fuel mixture while keeping it safely below self-ignition. Radial pistons placed in the rotor and mechanically timed to vanes motion, provide a more abrupt CR increase, which detonate the charge near an optimum rotor shaft angle. Resulting gases are then quickly expanded to minimize chamber wall heat losses throughout engine power stroke. This concept provides a useful torque range at a relatively lower RPM by taking advantage of the vane motor principle, where as many power strokes as rotor vanes occur per shaft revolution. Because there are so many, the combustion chambers can be sized much smaller than in a piston-crankshaft engine for a similar shaft power, which is a must for full-load HCCI. A simple and reliable mechanical timing between vane and piston motions is provided for accurate charge detonation control throughout engine load-speed map.
Unlike a piston-crankshaft, this vane-piston engine may be designed with very unequal CR and ER values:
An ER value much greater than CR followed by a cooled post-ignition recompression phase maximizes shaft power output by means of an unpublished thermodynamic cycle whose efficiency may surpass Atkinson cycle's to benefit power generators, surface vehicles or subsonic aircraft applications. A CR
greater than ER provides a high-pressure and high-frequency (over a hundred Hertz) pulse-generator at exhaust ports to control pulse detonation engines (PDE) for high-speed aerospace propulsion.
Controlling such a fast combustion in conventional piston-crankshaft engines throughout a useful load-speed range has proven difficult so far, as combustion chambers are sized for deflagration on SI, or flame diffusion on CI or Diesel, but not for detonation. In four-stroke engines two crankshaft revolutions are required for each power stroke, and the shallow sinusoidal compression ratio (CR) variation inherent to the piston-crank kinematics near top dead center (TDC) tends to cause either misfiring when engine is cold or runs too lean (CR value insufficient for detonation near TDC) or severe knocking, excessive vehicle noise-vibration-harshness (NVH) and structurally intolerable pressure rise rate during transients, when engine gets hot or runs higher load at lower revolutions per minute (RPM) (premature detonation problem). As a result, most HCCI
concepts in the works are based on sophisticated sensor-computer-actuator piston-crankshaft engines that run on HCCI mode at part load or idle, and revert to the century-old thermally and environmentally less efficient, but structurally gentler, SI or CI mode at full load, or on cold start.
This invention aims to control HCCI under all operating conditions, thus to minimize engine air emissions and heat losses by means of a vane-piston engine: A radial vane pump with vanes positively guided by inner and outer cams, compresses gradually air, or a homogeneous air-fuel mixture while keeping it safely below self-ignition. Radial pistons placed in the rotor and mechanically timed to vanes motion, provide a more abrupt CR increase, which detonate the charge near an optimum rotor shaft angle. Resulting gases are then quickly expanded to minimize chamber wall heat losses throughout engine power stroke. This concept provides a useful torque range at a relatively lower RPM by taking advantage of the vane motor principle, where as many power strokes as rotor vanes occur per shaft revolution. Because there are so many, the combustion chambers can be sized much smaller than in a piston-crankshaft engine for a similar shaft power, which is a must for full-load HCCI. A simple and reliable mechanical timing between vane and piston motions is provided for accurate charge detonation control throughout engine load-speed map.
Unlike a piston-crankshaft, this vane-piston engine may be designed with very unequal CR and ER values:
An ER value much greater than CR followed by a cooled post-ignition recompression phase maximizes shaft power output by means of an unpublished thermodynamic cycle whose efficiency may surpass Atkinson cycle's to benefit power generators, surface vehicles or subsonic aircraft applications. A CR
greater than ER provides a high-pressure and high-frequency (over a hundred Hertz) pulse-generator at exhaust ports to control pulse detonation engines (PDE) for high-speed aerospace propulsion.
Description
SPECIFICATION
Summary This invention relates to an internal combustion engine that allows Homogeneous Charge Compression Ignition (HCCI) operation suitable for stationary powerplant, surface or aerospace propulsion applications. Air emissions and heat losses in the engine thermodynamic cycle are significantly reduced thanks to a vane compressor-expander mechanically timed with radial pistons placed between the rotor vane slots to control charge detonation throughout the engine load-speed map. The combustion chambers may be easily designed with different compression and expansion ratios either to follow a new thermodynamic cycle that has the potential to surpass Atkinson cycle efficiency for power generator, surface vehicles and subsonic propeller applications where maximum shaft power is desirable, or for high-speed aerospace propulsion PDE application where mechanical work produced is sufficient to sustain shaft revolution and the balance of energy at exhaust can be used for high-pressure and high-frequency pulsing.
Background Many concepts were found from preliminary search in Canadian, US, European and Australian patent databases using keywords like "rotary engine, vane, HCCI". Some examples are listed below:
CA 1106291 Rotary internal combustion engine with paired edge seals In this concept, a vane pump is used like in the present application. However, the intent is only to increase the expansion phase in regards to the compression phase (Atkinson cycle), not to control charge detonation using a vane-piston compressor-expander like the present application does.
CA 2289136 HCCI internal combustion engine This concept is, in regards to its intent, probably the closest to the present application, as it provides a mechanical timing between a 1 and a 2 d stage compressor in the attempt to control detonation of a homogeneous charge. However, unlike the present application that uses a rotary compressor with many reduced size combustion chambers to achieve maximum torque at low RPM, this concept is based on a reciprocating 1st stage compressor, which requires one relatively larger chamber and a higher RPM to reach useful torque and power. This makes HCCI mode more difficult to control and, as shown in the representative drawing of the patent, requires spark ignition to start and to operate the engine at full load.
CA 2511267 Rotary engine with pivoting blades This concept, also known as "Quasiturbine" uses, like the present application, a rotary compressor having a non-sinusoidal CR. However, this compressor uses a rotor with four blades assembled to a deformable parallelogram, which limits the number of power strokes to four per revolution. The present application is based on a one-piece slotted rotor offering as many strokes per revolution as there are vanes placed in it.
CA 2027958 Rotary internal combustion engine This discloses another engine based on the vane pump design, whereas the present application discloses a vane-piston compressor.
US 7059,294B2, US 5,222,463 known as "Roundengine","rotoblock" and "MYTengine"
These engines that have pistons traveling along an annular cylinder by constant velocity or oscillating motion are unrelated to the proposed vane-piston motor principle.
CA 2085187,2208873, 2496157,2183527,2180198 Axial vane rotary engine... with slidable vane support... split vanes... sealing system... continuous fuel injection These disclose a CI or Diesel vane motor named "Radmax" that, like the present application, guides vane motion with cams in the stator inner flanges.
However, no pistons are included in this vane motor, and the vane sliding motion relative to the rotor is axial rather than radial, which presents a lower mechanical advantage (less vane lever arm about shaft axis) when compared to a radial vane concept.
CA 202671 (dated 1920) US 3,863,611 (1975) and US 6,554,596 (2003) confirm that a radial vane motor with vanes positively guided with outer and inner cams is certainly not a new idea.
The basis of the proposed invention resides in the use of outer and inner cams positively guided vanes, and pistons, to provide a multitude of small combustion chambers with a non-sinusoidal compression ratio to suit HCCI operation.
Summary This invention relates to an internal combustion engine that allows Homogeneous Charge Compression Ignition (HCCI) operation suitable for stationary powerplant, surface or aerospace propulsion applications. Air emissions and heat losses in the engine thermodynamic cycle are significantly reduced thanks to a vane compressor-expander mechanically timed with radial pistons placed between the rotor vane slots to control charge detonation throughout the engine load-speed map. The combustion chambers may be easily designed with different compression and expansion ratios either to follow a new thermodynamic cycle that has the potential to surpass Atkinson cycle efficiency for power generator, surface vehicles and subsonic propeller applications where maximum shaft power is desirable, or for high-speed aerospace propulsion PDE application where mechanical work produced is sufficient to sustain shaft revolution and the balance of energy at exhaust can be used for high-pressure and high-frequency pulsing.
Background Many concepts were found from preliminary search in Canadian, US, European and Australian patent databases using keywords like "rotary engine, vane, HCCI". Some examples are listed below:
CA 1106291 Rotary internal combustion engine with paired edge seals In this concept, a vane pump is used like in the present application. However, the intent is only to increase the expansion phase in regards to the compression phase (Atkinson cycle), not to control charge detonation using a vane-piston compressor-expander like the present application does.
CA 2289136 HCCI internal combustion engine This concept is, in regards to its intent, probably the closest to the present application, as it provides a mechanical timing between a 1 and a 2 d stage compressor in the attempt to control detonation of a homogeneous charge. However, unlike the present application that uses a rotary compressor with many reduced size combustion chambers to achieve maximum torque at low RPM, this concept is based on a reciprocating 1st stage compressor, which requires one relatively larger chamber and a higher RPM to reach useful torque and power. This makes HCCI mode more difficult to control and, as shown in the representative drawing of the patent, requires spark ignition to start and to operate the engine at full load.
CA 2511267 Rotary engine with pivoting blades This concept, also known as "Quasiturbine" uses, like the present application, a rotary compressor having a non-sinusoidal CR. However, this compressor uses a rotor with four blades assembled to a deformable parallelogram, which limits the number of power strokes to four per revolution. The present application is based on a one-piece slotted rotor offering as many strokes per revolution as there are vanes placed in it.
CA 2027958 Rotary internal combustion engine This discloses another engine based on the vane pump design, whereas the present application discloses a vane-piston compressor.
US 7059,294B2, US 5,222,463 known as "Roundengine","rotoblock" and "MYTengine"
These engines that have pistons traveling along an annular cylinder by constant velocity or oscillating motion are unrelated to the proposed vane-piston motor principle.
CA 2085187,2208873, 2496157,2183527,2180198 Axial vane rotary engine... with slidable vane support... split vanes... sealing system... continuous fuel injection These disclose a CI or Diesel vane motor named "Radmax" that, like the present application, guides vane motion with cams in the stator inner flanges.
However, no pistons are included in this vane motor, and the vane sliding motion relative to the rotor is axial rather than radial, which presents a lower mechanical advantage (less vane lever arm about shaft axis) when compared to a radial vane concept.
CA 202671 (dated 1920) US 3,863,611 (1975) and US 6,554,596 (2003) confirm that a radial vane motor with vanes positively guided with outer and inner cams is certainly not a new idea.
The basis of the proposed invention resides in the use of outer and inner cams positively guided vanes, and pistons, to provide a multitude of small combustion chambers with a non-sinusoidal compression ratio to suit HCCI operation.
-2-Brief Description of Drawings FIG 1 is a perspective view of the HCCI engine showing the key components of a design configuration with vanes and pistons placed in the rotor where four engine strokes are achieved within one full rotor shaft revolution.
FIG 2a is a front view of a particular design where the rotor has twelve vanes and pistons, and dissymmetric stator compression and expansion lobes to suit the Atkinson thermodynamic cycle.
FIG 2b is a simulation plot of the geometric compression ratio (CR) versus shaft angle of the design shown in FIG 2a FIG 2c is a view showing some design details pertaining to the vane-piston mechanical timing of the proposed engine design.
FIG 2d is a view showing a particular design where the rotor has twelve vanes and pistons, and where two engine strokes are achieved within one full rotor shaft revolution.
FIG 3a is a front view of a HCCI engine showing key components of a design configuration with a relatively large number of vanes and a stator equipped with two diametrically opposed ignition sectors, where all engine strokes are achieved in half a rotor shaft revolution.
FIG 3b is a front view of a HCCI engine showing some key components of a design configuration with a relatively large number of vanes and a stator equipped with three angularly equally spaced ignition sectors, where the engine four strokes are achieved within one third rotor shaft revolution, and where the expansion angular sectors are smaller than the compression angular sectors to maximize high pressure pulsing at exhaust and where the reduced expansion strokes remain sufficient to sustain engine revolution.
FIG 3c is a plot showing the pressure pulse output expected at the exhaust ports of design shown in FIG 3b.
FIG 3d is a side view of a PDE engine when timed by a high-pressure pulse generator like the one described in FIGs 3b, 3c to suit supersonic or hypersonic aerospace propulsion.
FIG 4 presents the thermodynamic cycle of the proposed engine in a temperature-entropy (T-S) diagram. The other graph compares the proposed concept with the conventional Otto (gasoline) and Diesel engines in regards to thermal efficiency.
FIG 5 is a comparative plot showing typical torque performance of conventional gasoline, Diesel engine, and the proposed design.
FIG 6a & 6b depict the thermal efficiency gains that the proposed vane-piston engine may achieve over a piston-crankshaft engine, when both engines run the HCCI mode.
FIG 7 is a T-S diagram based summary of all possible derivatives that the proposed engine concept offers, from PDE pressure pulse generator application to a very-high efficiency engine able to follow a new ideal thermodynamic cycle that approaches ideal Carnot cycle efficiency.
FIG 2a is a front view of a particular design where the rotor has twelve vanes and pistons, and dissymmetric stator compression and expansion lobes to suit the Atkinson thermodynamic cycle.
FIG 2b is a simulation plot of the geometric compression ratio (CR) versus shaft angle of the design shown in FIG 2a FIG 2c is a view showing some design details pertaining to the vane-piston mechanical timing of the proposed engine design.
FIG 2d is a view showing a particular design where the rotor has twelve vanes and pistons, and where two engine strokes are achieved within one full rotor shaft revolution.
FIG 3a is a front view of a HCCI engine showing key components of a design configuration with a relatively large number of vanes and a stator equipped with two diametrically opposed ignition sectors, where all engine strokes are achieved in half a rotor shaft revolution.
FIG 3b is a front view of a HCCI engine showing some key components of a design configuration with a relatively large number of vanes and a stator equipped with three angularly equally spaced ignition sectors, where the engine four strokes are achieved within one third rotor shaft revolution, and where the expansion angular sectors are smaller than the compression angular sectors to maximize high pressure pulsing at exhaust and where the reduced expansion strokes remain sufficient to sustain engine revolution.
FIG 3c is a plot showing the pressure pulse output expected at the exhaust ports of design shown in FIG 3b.
FIG 3d is a side view of a PDE engine when timed by a high-pressure pulse generator like the one described in FIGs 3b, 3c to suit supersonic or hypersonic aerospace propulsion.
FIG 4 presents the thermodynamic cycle of the proposed engine in a temperature-entropy (T-S) diagram. The other graph compares the proposed concept with the conventional Otto (gasoline) and Diesel engines in regards to thermal efficiency.
FIG 5 is a comparative plot showing typical torque performance of conventional gasoline, Diesel engine, and the proposed design.
FIG 6a & 6b depict the thermal efficiency gains that the proposed vane-piston engine may achieve over a piston-crankshaft engine, when both engines run the HCCI mode.
FIG 7 is a T-S diagram based summary of all possible derivatives that the proposed engine concept offers, from PDE pressure pulse generator application to a very-high efficiency engine able to follow a new ideal thermodynamic cycle that approaches ideal Carnot cycle efficiency.
-3-Description of Preferred Embodiment Referring to Figure 1, an engine, with intake, exhaust, cooling, sealing and lubrication systems not entirely displayed, is composed of a stator (8) closed at one end by flange (7) with bolts through circumferential holes like (5) and (9) and closed at the other end by an opposite flange not shown. This assembly encloses a rotor (20) coupled to output shaft (26) placed in bearing holes like (3). Rotor (20) is equipped with multiple slots (21) that are oriented in radial directions and are spaced equally. Each slot (21) guides a vane (22) along the rotor radius. Each vane (22) is also guided on its outer end by housing (10) of stator (8) and on its inner end by inner cam (33) installed on the stator ends like (7). The slotted rotor (20) is structurally reinforced with collar (11) on one side and collar (23) on the other bolted through holes like (13), (18) and (25). These collars are also used as provision for sealing purposes. Rotor (20) is also equipped with pistons like piston (15) recessed in (14) between two consecutive vanes. Each piston is driven along rotor radial direction by means of one pin (16) installed through hole (27), through rotor oblong hole (17) and through slots (12) and (24) of rotor collars (11) and (23). Each pin (16) is fitted with bearings (28), (29) at both ends to ride on cams (34), (35) of stator ends (7) as the rotor rotates, and serves to provide mechanical timing between the rotary motion of rotor (20) and the reciprocating motion of radial pistons like (15). A recess (19) is provided on each side of rotor (20) to avoid interference with the vane inner cam (33) of ends (7).
Air, or an homogeneous charge, is naturally aspirated, supercharged, or turbo-charged, during the intake phase through intake port (6). As an alternative to a carbureted or port-injected mixture, fuel is mixed to air through direct injector nozzle (38). The charge is then compressed below detonation point until Top Dead Center (TDC) by vane pumping effect. The charge is then suddenly further compressed above detonation point under the effect of piston (15) extending motion. The timing angle (c) is set for optimum conditions.
Piston (15) then suddenly retreats to provide rapid adiabatic cooling of the detonated gas, which keeps expanding at a lower rate for the remainder of the power stroke. As a last phase, the expanded gas exhausts through port (4).
Such a vane-piston arrangement allows the engine to run full load at near-stoichiometric equivalence ratios and may be set to accommodate fuels of various octane numbers, like natural gas or hydrogen, where CR values may approach Diesel levels during detonation for the sake of thermal efficiency.
Pressure rise rate, NVH and heat losses caused by maximum charge temperature and pressure due to high equivalence ratio and detonation CR are mitigated by the small combustion chambers size and the rapid compression and post-detonation expansion features that this design offers. A desirable load sharing occurs between rotor axle bearings (3) and piston cam followers (28) under the effect of detonation pressure forces. When each cam driven radial piston triggers detonation, the chamber bounding vanes are almost entirely retreated in their respective slots with little or no radial sliding motion. These vanes are thus protected from excessive bending, friction heat losses and wear due to the tremendous pressure rise that accompanies detonation. Unlike conventional piston-crankshaft that exhibit fairly large combustion chambers for a given engine capacity, a cam-driven-vane-piston can be easily arranged as a multitude of combustion chambers better sized to accommodate full-load HCCI
operation.
Air, or an homogeneous charge, is naturally aspirated, supercharged, or turbo-charged, during the intake phase through intake port (6). As an alternative to a carbureted or port-injected mixture, fuel is mixed to air through direct injector nozzle (38). The charge is then compressed below detonation point until Top Dead Center (TDC) by vane pumping effect. The charge is then suddenly further compressed above detonation point under the effect of piston (15) extending motion. The timing angle (c) is set for optimum conditions.
Piston (15) then suddenly retreats to provide rapid adiabatic cooling of the detonated gas, which keeps expanding at a lower rate for the remainder of the power stroke. As a last phase, the expanded gas exhausts through port (4).
Such a vane-piston arrangement allows the engine to run full load at near-stoichiometric equivalence ratios and may be set to accommodate fuels of various octane numbers, like natural gas or hydrogen, where CR values may approach Diesel levels during detonation for the sake of thermal efficiency.
Pressure rise rate, NVH and heat losses caused by maximum charge temperature and pressure due to high equivalence ratio and detonation CR are mitigated by the small combustion chambers size and the rapid compression and post-detonation expansion features that this design offers. A desirable load sharing occurs between rotor axle bearings (3) and piston cam followers (28) under the effect of detonation pressure forces. When each cam driven radial piston triggers detonation, the chamber bounding vanes are almost entirely retreated in their respective slots with little or no radial sliding motion. These vanes are thus protected from excessive bending, friction heat losses and wear due to the tremendous pressure rise that accompanies detonation. Unlike conventional piston-crankshaft that exhibit fairly large combustion chambers for a given engine capacity, a cam-driven-vane-piston can be easily arranged as a multitude of combustion chambers better sized to accommodate full-load HCCI
operation.
-4-Referring to Figure 2a, engine rotor (20) is shown installed in a geometrically dissymmetric stator housing (10) that provides an expansion chamber (30) with a volume greater than the volume of compression chamber (31). Each vane (removed in the Figure) slides in rotor radial slot (21), and is positively guided through stator housing (10) and stator end inner cam (33) to constitute the vane compressor-expander of the engine. Piston compression-expansion phase is controlled by means of radial pistons, not shown in the Figure, placed on the rotor cylindrical surface between vane slots (21).
Piston reciprocating motion is controlled by means of followers and cams (34), (35) placed in the stator flanges. As the engine runs, piston positive pressure forces are reacted by inner cam (35) and centrifugal and negative pressure forces are reacted by outer cam (34).
Angular positioning of profile (36) of the inner and outer cams (35), (34) relative to vane compressor TDC
provides mechanical timing between the vanes and pistons. Although not shown in the Figure, depending on engine thermal allowance, the descending ramp of piston cam wave shape (36) may be delayed relative to its climbing ramp to coincide with the angular pitch of two successive vanes. The resulting driving and resisting reactions of the piston cam followers will then tend to cancel each other when detonation occurs to further reduce NVH and enhance smoother load/RPM transients. The particular case of a four-stroke engine is shown in this figure, where local shaping in angular sector (39) of vane stator housing (10) and vane inner cam (33) provides exhaust gas scavenging and charge intake.
Referring to Figures 2b, a plot simulation of the CR, ER variations versus shaft angle of engine described in Figure 2a is displayed in solid line. The dashed portion shows the CR, ER variations that the engine would exhibit without the effect of the radial piston compressor/expander. It may be appreciated from these curves that a proper setting of the timing between vanes and pistons allows to reach a CR variation that would suit HCCI operation without relying on any sophisticated sensor-computer-actuator engine technology.
Referring to Figure 2c, engine rotor (20) is shown with slot structural stiffeners (42) and with the vanes removed from slots (21) for the sake of clarity. One piston (15) is shown installed in retreated position with cross pin (16) mounted with journal bearing cam followers (28), (29) riding on their respective piston cam profiles (34), (35). Piston (15) is guided at its inner extremity by means of stem (40) to help react circumferential loads induced by journal bearings (28) (29) riding over ramps in cam profiles (34), (35). The view shows also in dashed lines the collars (11) (23) and their attachment holes (18). To compensate for loss in chamber volume-to-area ratio and maintain stable detonation control on smaller size engines, piston head (15) shown flat in the figure may need to be designed with a concave shape.
Referring to Figure 2d, engine rotor (20) is shown installed in stator housing (10) that allows each combustion chamber (30) to describe a two-stroke cycle. Each vane (22) slides in rotor radial slot (21), and is positively guided between circular stator housing (10) and stator end inner cam (33) to constitute the vane compressor-expander of the engine. Piston compression-expansion phase is controlled by means of radial pistons, not shown in the Figure, placed on the rotor cylindrical surface between vane slots (21) with followers riding on piston inner and outer cams not shown in the Figure. An optimized angular overlap (39) between exhaust port (4) located on one stator flange and intake port (6) located on opposite stator flange causes the exhausting gases to aspirate intake air.
Then, high-pressure direct fuel injection through nozzle (38) provides uniform charge mixing. As load on engine shaft (26) rises, the required increased amount of fuel injected to maintain RPM cools the charge further, thus compensates rise in equivalence ratio and contributes to full-load HCCI stability of operation. Recess (19) is provided on each side of rotor (20) to avoid interference with vane inner cams (33).
Piston reciprocating motion is controlled by means of followers and cams (34), (35) placed in the stator flanges. As the engine runs, piston positive pressure forces are reacted by inner cam (35) and centrifugal and negative pressure forces are reacted by outer cam (34).
Angular positioning of profile (36) of the inner and outer cams (35), (34) relative to vane compressor TDC
provides mechanical timing between the vanes and pistons. Although not shown in the Figure, depending on engine thermal allowance, the descending ramp of piston cam wave shape (36) may be delayed relative to its climbing ramp to coincide with the angular pitch of two successive vanes. The resulting driving and resisting reactions of the piston cam followers will then tend to cancel each other when detonation occurs to further reduce NVH and enhance smoother load/RPM transients. The particular case of a four-stroke engine is shown in this figure, where local shaping in angular sector (39) of vane stator housing (10) and vane inner cam (33) provides exhaust gas scavenging and charge intake.
Referring to Figures 2b, a plot simulation of the CR, ER variations versus shaft angle of engine described in Figure 2a is displayed in solid line. The dashed portion shows the CR, ER variations that the engine would exhibit without the effect of the radial piston compressor/expander. It may be appreciated from these curves that a proper setting of the timing between vanes and pistons allows to reach a CR variation that would suit HCCI operation without relying on any sophisticated sensor-computer-actuator engine technology.
Referring to Figure 2c, engine rotor (20) is shown with slot structural stiffeners (42) and with the vanes removed from slots (21) for the sake of clarity. One piston (15) is shown installed in retreated position with cross pin (16) mounted with journal bearing cam followers (28), (29) riding on their respective piston cam profiles (34), (35). Piston (15) is guided at its inner extremity by means of stem (40) to help react circumferential loads induced by journal bearings (28) (29) riding over ramps in cam profiles (34), (35). The view shows also in dashed lines the collars (11) (23) and their attachment holes (18). To compensate for loss in chamber volume-to-area ratio and maintain stable detonation control on smaller size engines, piston head (15) shown flat in the figure may need to be designed with a concave shape.
Referring to Figure 2d, engine rotor (20) is shown installed in stator housing (10) that allows each combustion chamber (30) to describe a two-stroke cycle. Each vane (22) slides in rotor radial slot (21), and is positively guided between circular stator housing (10) and stator end inner cam (33) to constitute the vane compressor-expander of the engine. Piston compression-expansion phase is controlled by means of radial pistons, not shown in the Figure, placed on the rotor cylindrical surface between vane slots (21) with followers riding on piston inner and outer cams not shown in the Figure. An optimized angular overlap (39) between exhaust port (4) located on one stator flange and intake port (6) located on opposite stator flange causes the exhausting gases to aspirate intake air.
Then, high-pressure direct fuel injection through nozzle (38) provides uniform charge mixing. As load on engine shaft (26) rises, the required increased amount of fuel injected to maintain RPM cools the charge further, thus compensates rise in equivalence ratio and contributes to full-load HCCI stability of operation. Recess (19) is provided on each side of rotor (20) to avoid interference with vane inner cams (33).
-5-Referring to Figure 3a, an engine, with intake, exhaust, cooling, sealing and lubrication systems not displayed, encloses a rotor (38) equipped with vanes, not shown for clarity, guided on their outer end by stator housing (32) and on their inner end by inner cam (33) installed on the stator ends. Rotor recess (19) shown on rotor (38) is provided to prevent interference with vane inner cam (33). A homogeneous charge is naturally aspirated, supercharged, or turbo-charged, during the intake phase represented by angular sector (a-b) through intake port (6). The charge is then compressed below detonation point near vane compressor TDC under the vane pump effect during angular sector (b-c).
The charge is then suddenly further compressed during sector (c-d) above detonation point under the effect of the radial expansion of pistons not shown for clarity. The detonated gas then expands throughout the power stroke of the engine during sector (d-e) until it exhausts through port (4) during sector (e-a). This Figure depicts a possible configuration of the proposed vane-piston motor concept, where two diametrically opposed ignition sectors are provided rather than one to benefit rotors having a larger number of vanes for lower RPM operation and provide a more uniform stator temperature gradient which helps maintain high volumetric efficiency by minimizing internal gas leaks due to relative thermal expansion between stator and moving rotor and vanes. Detonations may be phased simultaneously between the two ignition angular sectors to provide a desirable stress relief on the rotor shaft bearings. A rotor with an odd number of vanes causes detonations at the two ignition sectors to be interlaced, which may help further reduce low RPM NVH.
Referring to Figure 3b, an engine, with intake, exhaust, cooling, sealing and lubrication systems not displayed, is shown with stator (8) having a geometrically dissymmetric housing that provides expansion chambers volume (30) smaller than compression chambers volume (31).
Stator (8) encloses a rotor (20) equipped with vanes, not shown for clarity, guided on their outer end by stator housing and on their inner end by inner cam (33) installed on the stator ends. Recess (19) shown on rotor (20) is provided to clear the vane inner cam (33). Slot structural stiffeners like (42) are also shown on rotor (20). Air is naturally aspirated, supercharged, or turbo-charged, during the intake phase represented by angular sector (a,-bl) through intake port (6). Fuel is directly injected at the beginning of the compression phase through nozzle (38). The charge is then compressed below detonation point near the vane compressor TDC under the vane pump effect during angular sector (b,-c,).
The charge is then suddenly further compressed above detonation point during sector (c,-d,), under radial expansion of the piston not shown for clarity. The detonated gas then expands throughout engine power stroke during sector (d,-e,) until it exhausts through port (4) during sector (e,-a2). This Figure depicts a possible configuration of the proposed vane-piston motor concept, where three equally spaced ignition sectors are provided rather than one to benefit rotors having a larger number of vanes for lower RPM operation and also to provide a more uniform stator temperature gradient to help maintain a high volumetric efficiency by minimizing gas leaks due to relative thermal expansion between the stator and the moving rotor and vanes. A multiple of three vanes in the rotor allow detonations to be phased simultaneously between the three ignition sectors, which provide a desirable stress relief on the rotor shaft bearings. A
rotor with a number of vanes not being a multiple of three causes detonations between the three ignition sectors to be interlaced, which may help further reduce low RPM NVH. Such configurations provide high pressure pulses with frequencies that may suit Pulse Detonation Engine (PDE) applications.
The charge is then suddenly further compressed during sector (c-d) above detonation point under the effect of the radial expansion of pistons not shown for clarity. The detonated gas then expands throughout the power stroke of the engine during sector (d-e) until it exhausts through port (4) during sector (e-a). This Figure depicts a possible configuration of the proposed vane-piston motor concept, where two diametrically opposed ignition sectors are provided rather than one to benefit rotors having a larger number of vanes for lower RPM operation and provide a more uniform stator temperature gradient which helps maintain high volumetric efficiency by minimizing internal gas leaks due to relative thermal expansion between stator and moving rotor and vanes. Detonations may be phased simultaneously between the two ignition angular sectors to provide a desirable stress relief on the rotor shaft bearings. A rotor with an odd number of vanes causes detonations at the two ignition sectors to be interlaced, which may help further reduce low RPM NVH.
Referring to Figure 3b, an engine, with intake, exhaust, cooling, sealing and lubrication systems not displayed, is shown with stator (8) having a geometrically dissymmetric housing that provides expansion chambers volume (30) smaller than compression chambers volume (31).
Stator (8) encloses a rotor (20) equipped with vanes, not shown for clarity, guided on their outer end by stator housing and on their inner end by inner cam (33) installed on the stator ends. Recess (19) shown on rotor (20) is provided to clear the vane inner cam (33). Slot structural stiffeners like (42) are also shown on rotor (20). Air is naturally aspirated, supercharged, or turbo-charged, during the intake phase represented by angular sector (a,-bl) through intake port (6). Fuel is directly injected at the beginning of the compression phase through nozzle (38). The charge is then compressed below detonation point near the vane compressor TDC under the vane pump effect during angular sector (b,-c,).
The charge is then suddenly further compressed above detonation point during sector (c,-d,), under radial expansion of the piston not shown for clarity. The detonated gas then expands throughout engine power stroke during sector (d,-e,) until it exhausts through port (4) during sector (e,-a2). This Figure depicts a possible configuration of the proposed vane-piston motor concept, where three equally spaced ignition sectors are provided rather than one to benefit rotors having a larger number of vanes for lower RPM operation and also to provide a more uniform stator temperature gradient to help maintain a high volumetric efficiency by minimizing gas leaks due to relative thermal expansion between the stator and the moving rotor and vanes. A multiple of three vanes in the rotor allow detonations to be phased simultaneously between the three ignition sectors, which provide a desirable stress relief on the rotor shaft bearings. A
rotor with a number of vanes not being a multiple of three causes detonations between the three ignition sectors to be interlaced, which may help further reduce low RPM NVH. Such configurations provide high pressure pulses with frequencies that may suit Pulse Detonation Engine (PDE) applications.
-6-Figure 3c, shows a time plot of a pressure pulsing that may be achieved for a relatively low shaft rotation speed in the particular configuration of the engine design shown in Figure 3b. Assuming a rotation speed of 600 RPM, the resulting pulse frequency would be 600 RPM/60sec* 12vanes = 120Hz.
Interlacing the detonation events would raise the pulsing frequency to 3 times 120 = 360Hz. On Figure 3c (Pa) is the ambient pressure after chamber blow-down, and (PId) represents a pressure level that would be required to control detonations in a PDE initiator chamber located downstream.
Referring to Figure 3d, an engine like the one disclosed in Figure 3b is shown as unit (8) when integrated into a valve-less pulse detonation engine (PDE) for high-speed aerospace propulsion operation. Air enters intake ports (6) of unit (8), undergoes the HCCI
combustion process described above, before expanding into PDE initiator chamber (53) by means of high-pressure pulses through exhaust ports (4) as plotted in Figure 3c. Air also enters the PDE through the main annular intake where a bank of fuel injectors is placed to mix the charge. Part of the charge is transferred to initiator chamber (53) by means of intake ports (56). Additional fuel and oxygen injectors (52) may be used to raise charge detonation sensitivity to pressure pulsing. Charge detonation takes place in chamber (53) driven by the high pressure pulsing of unit (8). Resulting detonation shock wave propagates through initiator nozzle (57) and wave diffraction occurs to main engine chamber (58), which causes shock wave (55), whose pressure variations are plotted on the graph below, to generate thrust by propagating through main engine nozzle (59).
Interlacing the detonation events would raise the pulsing frequency to 3 times 120 = 360Hz. On Figure 3c (Pa) is the ambient pressure after chamber blow-down, and (PId) represents a pressure level that would be required to control detonations in a PDE initiator chamber located downstream.
Referring to Figure 3d, an engine like the one disclosed in Figure 3b is shown as unit (8) when integrated into a valve-less pulse detonation engine (PDE) for high-speed aerospace propulsion operation. Air enters intake ports (6) of unit (8), undergoes the HCCI
combustion process described above, before expanding into PDE initiator chamber (53) by means of high-pressure pulses through exhaust ports (4) as plotted in Figure 3c. Air also enters the PDE through the main annular intake where a bank of fuel injectors is placed to mix the charge. Part of the charge is transferred to initiator chamber (53) by means of intake ports (56). Additional fuel and oxygen injectors (52) may be used to raise charge detonation sensitivity to pressure pulsing. Charge detonation takes place in chamber (53) driven by the high pressure pulsing of unit (8). Resulting detonation shock wave propagates through initiator nozzle (57) and wave diffraction occurs to main engine chamber (58), which causes shock wave (55), whose pressure variations are plotted on the graph below, to generate thrust by propagating through main engine nozzle (59).
-7-Figure 4 left graph depicts the thermodynamic cycle in a Temperature-Entropy (T-S) diagram of the proposed design concept. The 1-2 evolution is a gradual adiabatic compression from ambient conditions of the vane pump (called 1 sC stage) limited to a CR below the mixture detonation point to approach the vane pump TDC. The 2-3 evolution is the additional and more rapid adiabatic compression provided by the radial pistons (called 2nd stage) that causes detonation of the homogeneous charge within a few shaft angle degrees. The ensuing 3-4 sequence is the quasi-immediate temperature rise resulting from detonation, which may be represented by an isometric, or constant volume evolution.
The CR and mixture equivalence ratio are so designed to limit the maximum temperature (Tmax) and pressure to a technologically acceptable value under full load condition. The 4-5 evolution is the expansion provided by the 2"d stage of the engine: when radial pistons retreat towards the rotor center.
The shape of the piston cam located on each stator inner flange imposes this motion in order to control an expansion rapid enough to minimize unwanted heat exchange with the chamber walls. The 5-6 evolution is the continuation of the adiabatic expansion provided by the 1St stage and completes the engine power stroke. 6-1 is, for the case of an Otto cycle, the non-adiabatic portion of the power stroke and includes also the exhaust and intake strokes. The proposed engine may be designed with a geometrically dissymmetric stator yielding an expansion ratio value greater than the CR value, which modifies the Otto cycle into a more thermally efficient Atkinson cycle for high load conditions. As a result, the power stroke lasts longer and the end point 6 moves down to point 7.
Figure 4 right graph compares in a T-S diagram the fundamental thermodynamic differences between the proposed engine concept, the Otto cycle (spark ignition and deflagration of homogeneous gasoline mixture) and the Diesel cycle (compression-ignition and flame diffusion).
Assuming that all three cycles are limited by the same (Tmax) value, their idealized respective thermal efficiencies may be readily computed from the graph as follows:
If H=Area { a3' 4b }, amount of heat required by the proposed concept, and W=Area 113'461, (W=Area 113'47) for an Atkinson cycle), amount of mechanical work produced, the thermal efficiency of the cycle followed by the proposed concept is W/H.
If Hd=Area { a3"8c }, amount of heat required by the Diesel engine, and Wd=Area { 13"89), amount of mechanical work produced, Diesel cycle thermal efficiency is Wd/Hd.
If Hg=Area { a28c }, amount of heat required by the Otto cycle and Wg= {1289 }, amount of mechanical work produced, the ideal gasoline engine, thermal efficiency is Wg/Hg. It should be noted that combustion segment {28} is a relatively slow process limited by deflagration speed where heat losses with chamber walls are substantial. For that reason, a gasoline engine is unable to follow a true Otto cycle, and point { 21 should be placed somewhat lower than shown on the graph, which further reduces gasoline engine thermal efficiency.
By inspection of the graph, the known fact that Diesel cycle efficiency is higher than the true Otto cycle efficiency is verified: (Wd/Hd) > (Wg/Hg), because of the higher CR that Diesel is able to achieve when combustion begins. It also appears that efficiency offered by the proposed HCCI concept exceeds Diesel cycle efficiency: (W/H) > (Wd/Hd), because its rapid combustion process allows a true constant volume rather than a constant pressure evolution, (HCCI combustion segment {3'4} occurs much more rapidly than any of the gasoline deflagration 128) or Diesel flame diffusion {
3"8 1 segments) which greatly reduces heat losses.
It is suggested that the thermodynamic cycle involved in the proposed engine concept be described as a true Otto, or Atkinson cycle having the ability to approach Diesel compression ratios before ignition.
The CR and mixture equivalence ratio are so designed to limit the maximum temperature (Tmax) and pressure to a technologically acceptable value under full load condition. The 4-5 evolution is the expansion provided by the 2"d stage of the engine: when radial pistons retreat towards the rotor center.
The shape of the piston cam located on each stator inner flange imposes this motion in order to control an expansion rapid enough to minimize unwanted heat exchange with the chamber walls. The 5-6 evolution is the continuation of the adiabatic expansion provided by the 1St stage and completes the engine power stroke. 6-1 is, for the case of an Otto cycle, the non-adiabatic portion of the power stroke and includes also the exhaust and intake strokes. The proposed engine may be designed with a geometrically dissymmetric stator yielding an expansion ratio value greater than the CR value, which modifies the Otto cycle into a more thermally efficient Atkinson cycle for high load conditions. As a result, the power stroke lasts longer and the end point 6 moves down to point 7.
Figure 4 right graph compares in a T-S diagram the fundamental thermodynamic differences between the proposed engine concept, the Otto cycle (spark ignition and deflagration of homogeneous gasoline mixture) and the Diesel cycle (compression-ignition and flame diffusion).
Assuming that all three cycles are limited by the same (Tmax) value, their idealized respective thermal efficiencies may be readily computed from the graph as follows:
If H=Area { a3' 4b }, amount of heat required by the proposed concept, and W=Area 113'461, (W=Area 113'47) for an Atkinson cycle), amount of mechanical work produced, the thermal efficiency of the cycle followed by the proposed concept is W/H.
If Hd=Area { a3"8c }, amount of heat required by the Diesel engine, and Wd=Area { 13"89), amount of mechanical work produced, Diesel cycle thermal efficiency is Wd/Hd.
If Hg=Area { a28c }, amount of heat required by the Otto cycle and Wg= {1289 }, amount of mechanical work produced, the ideal gasoline engine, thermal efficiency is Wg/Hg. It should be noted that combustion segment {28} is a relatively slow process limited by deflagration speed where heat losses with chamber walls are substantial. For that reason, a gasoline engine is unable to follow a true Otto cycle, and point { 21 should be placed somewhat lower than shown on the graph, which further reduces gasoline engine thermal efficiency.
By inspection of the graph, the known fact that Diesel cycle efficiency is higher than the true Otto cycle efficiency is verified: (Wd/Hd) > (Wg/Hg), because of the higher CR that Diesel is able to achieve when combustion begins. It also appears that efficiency offered by the proposed HCCI concept exceeds Diesel cycle efficiency: (W/H) > (Wd/Hd), because its rapid combustion process allows a true constant volume rather than a constant pressure evolution, (HCCI combustion segment {3'4} occurs much more rapidly than any of the gasoline deflagration 128) or Diesel flame diffusion {
3"8 1 segments) which greatly reduces heat losses.
It is suggested that the thermodynamic cycle involved in the proposed engine concept be described as a true Otto, or Atkinson cycle having the ability to approach Diesel compression ratios before ignition.
-8-Figure 5 graph presents a qualitative comparison of torque curves versus RPM
between the proposed engine design (1) and typical gasoline (3) and Diesel (2) piston engines, as well as the compressed-air vane motor (4) (known to be capable of maximum torque at zero RPM) that the vane compressor expander of the proposed design is derived from.
Figure 6a depicts in a T-S diagram the thermal efficiency loss that a conventional piston-crankshaft engine exhibits in a real thermodynamic cycle, where negative mechanical work area shown with a minus sign deducts to the useful work area shown with a plus sign. Negative area (3)(4)(5) shown on figure LHS is imposed by the sinusoidal CR shown on figure RHS: Detonation occurring at point (2) is so fast in regards to the piston reciprocating motion near TDC that the piston keeps compressing the detonated gases at a relatively slow rate which favors unwanted heat rejection.
Figure 6b depicts the thermal efficiency loss that the proposed vane-piston engine exhibits in a real thermodynamic cycle where negative mechanical work area shown with a minus sign deducts to the useful work area shown with a plus sign. Negative area (3)(4)(5) shown on figure LHS is reduced by the non-sinusoidal CR shown on figure RHS: Detonation occurring at point (2) is followed by the pistons rapid reciprocating motion which minimizes unwanted heat rejection.
Reciprocating motion under high post-ignition pressures is structurally achievable by these pistons thank to their reduced size.
Figure 7 summarizes in a T-S diagram all the possible derivatives that the proposed engine offers:
(A) shows a PDE pressure pulse generator application, where engine CR is made greater than ER, and engine is able to follow the (1)-(2)-(3)-(4)-(5) cycle.
(B) shows a configuration where CR = ER, and engine is able to follow an "Otto" cycle (1)-(2)-(3)-(4)-(6) with a constant volune heat rejection, or exhaust phase.
(C) shows a configuration where ER is made greater than CR, and engine is able to follow an "Atkinson" or "Miller" cycle (1)-(2)-(3)-(4)-(7) with a constant pressure heat rejection, or exhaust phase for an increased thermal efficiency.
(D) shows a configuration where ER is made even greater than CR, and with an additional constant volume and compression phase CR' connected to a cold source, and placed between the engine power stroke and the exhaust phase.
Engine (D) describes a "new" thermodynamic cycle (1)-(2)-(3)-(4)-(8)-(9)-(10), where (1)-(2) is an isentropic compression, (2)-(3) is an iso-volume heat addition caused by HCCI
combustion, (3)-(4) is the compression remainder occurring post-ignition having negligible effect on cycle as explained in Figure 6b, (4)-(8) is a power stroke or isentropic expansion ER caused by engine vanes and pistons, (8)-
between the proposed engine design (1) and typical gasoline (3) and Diesel (2) piston engines, as well as the compressed-air vane motor (4) (known to be capable of maximum torque at zero RPM) that the vane compressor expander of the proposed design is derived from.
Figure 6a depicts in a T-S diagram the thermal efficiency loss that a conventional piston-crankshaft engine exhibits in a real thermodynamic cycle, where negative mechanical work area shown with a minus sign deducts to the useful work area shown with a plus sign. Negative area (3)(4)(5) shown on figure LHS is imposed by the sinusoidal CR shown on figure RHS: Detonation occurring at point (2) is so fast in regards to the piston reciprocating motion near TDC that the piston keeps compressing the detonated gases at a relatively slow rate which favors unwanted heat rejection.
Figure 6b depicts the thermal efficiency loss that the proposed vane-piston engine exhibits in a real thermodynamic cycle where negative mechanical work area shown with a minus sign deducts to the useful work area shown with a plus sign. Negative area (3)(4)(5) shown on figure LHS is reduced by the non-sinusoidal CR shown on figure RHS: Detonation occurring at point (2) is followed by the pistons rapid reciprocating motion which minimizes unwanted heat rejection.
Reciprocating motion under high post-ignition pressures is structurally achievable by these pistons thank to their reduced size.
Figure 7 summarizes in a T-S diagram all the possible derivatives that the proposed engine offers:
(A) shows a PDE pressure pulse generator application, where engine CR is made greater than ER, and engine is able to follow the (1)-(2)-(3)-(4)-(5) cycle.
(B) shows a configuration where CR = ER, and engine is able to follow an "Otto" cycle (1)-(2)-(3)-(4)-(6) with a constant volune heat rejection, or exhaust phase.
(C) shows a configuration where ER is made greater than CR, and engine is able to follow an "Atkinson" or "Miller" cycle (1)-(2)-(3)-(4)-(7) with a constant pressure heat rejection, or exhaust phase for an increased thermal efficiency.
(D) shows a configuration where ER is made even greater than CR, and with an additional constant volume and compression phase CR' connected to a cold source, and placed between the engine power stroke and the exhaust phase.
Engine (D) describes a "new" thermodynamic cycle (1)-(2)-(3)-(4)-(8)-(9)-(10), where (1)-(2) is an isentropic compression, (2)-(3) is an iso-volume heat addition caused by HCCI
combustion, (3)-(4) is the compression remainder occurring post-ignition having negligible effect on cycle as explained in Figure 6b, (4)-(8) is a power stroke or isentropic expansion ER caused by engine vanes and pistons, (8)-
(9) is an iso-volume heat rejection, (9)-(10) is an isothermal compression CR' that pursues heat rejection, and (10)-(1) is an iso-pressure as shown, or iso-volume exhaust followed by an intake phase.
Due to geometric constraints of the physical engine, points (8) and (9) are likely to be distinct and to lie on an iso-volume curve corresponding to the engine maximum practical ER (ER, CR',,,ax=30-40).
However, this new cycle may be idealized, assuming that points (8), (9) coincide, and that points (10), (1) coincide, and that all these points lie on the engine intake temperature isothermal T1=300K.
Assuming also that points (3), (4) coincide, and lie on the T3=2700K
isothermal, by application of the isentropic evolution and ideal gas state equations for CR=12, the following thermal efficiencies are obtained: Otto=63%, Atkinson=71%, New=83%, and Carnot=89%. From these numbers, it may be appreciated that, by an adequate tailoring of the CR, CR' and ER values, the proposed engine design is able to describe a new ideal thermodynamic cycle: iso-S, iso-V, iso-S, iso-T
that may potentially approach the ideal Carnot cycle efficiency.
Due to geometric constraints of the physical engine, points (8) and (9) are likely to be distinct and to lie on an iso-volume curve corresponding to the engine maximum practical ER (ER, CR',,,ax=30-40).
However, this new cycle may be idealized, assuming that points (8), (9) coincide, and that points (10), (1) coincide, and that all these points lie on the engine intake temperature isothermal T1=300K.
Assuming also that points (3), (4) coincide, and lie on the T3=2700K
isothermal, by application of the isentropic evolution and ideal gas state equations for CR=12, the following thermal efficiencies are obtained: Otto=63%, Atkinson=71%, New=83%, and Carnot=89%. From these numbers, it may be appreciated that, by an adequate tailoring of the CR, CR' and ER values, the proposed engine design is able to describe a new ideal thermodynamic cycle: iso-S, iso-V, iso-S, iso-T
that may potentially approach the ideal Carnot cycle efficiency.
Claims (8)
1 A Homogeneous Charge Compression Ignition (HCCI) rotary engine composed of a slotted rotor that receives a multitude of radial vanes positively guided by inner and outer cams placed in the engine stator, and where the said cams have a vane retreating ramp throughout the pre-ignition compression angular sector until their near full retreat into the rotor slot followed by an ignition angular sector where vanes are kept in near retreated position, followed by an angular sector where vanes are extended outward during power stroke, and where one radial piston is placed on each angular sector defined by two successive rotor slots, and where the radial motion of said pistons is mechanically controlled by means of followers and additional cams having piston expanding and retreating ramps placed in one, or both, of the stator flanges.
2 An engine as defined in claim 1, where vane cams have a relatively shallow ramp throughout the compression and expansion angular sectors, and where piston cams have relatively steep ramps in the area of the ignition angular sector.
3 An engine as defined in any one of claims 1 or 2, where the pre-ignition compression stroke sector is preceded by two adjacent angular sectors where pistons are in near extended position, and vanes are almost entirely retreated, then extended again during exhaust and intake strokes of the engine.
4 An engine as defined in any one of claims 1, 2 or 3, where power stroke, or expansion volume, is greater than the pre-ignition compression volume.
An engine as defined in any one of claims 1, 2 or 3, where power stroke, or expansion volume, is smaller than the pre-ignition compression volume.
6 An engine as defined in any one of claims 1, 2, 3, 4 or 5, where the cam-driven vanes and pistons provide one pre-ignition compression stroke, one ignition phase, then one expansion or power stroke, within one full rotor shaft revolution.
7 An engine as defined in any one of claims 1, 2, 3, 4 or 5, where the cam-driven vanes and pistons provide one pre-ignition compression stroke, one ignition phase, then one expansion or power stroke within half, or any higher integer fraction of one full rotor shaft revolution.
8 An engine as defined in any one of claims 1, 2, 3, 4, 6 or 7 that exhibits an additional compression stroke sector, or post-ignition recompression stroke sector placed after the expansion or power stroke sector and before the engine exhaust and intake sectors, and where the said recompression stroke is cooled or accompanied with heat rejection.
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| WO2018020496A1 (en) * | 2016-07-25 | 2018-02-01 | Tenenbaum Yehonatan | Split-cycle engine with a variable displacement compressor and a rotary motor |
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| CN102226415B (en) * | 2011-05-06 | 2013-10-09 | 上海发电设备成套设计研究院 | Device and method for monitoring and controlling security risk of steam turbine rotor in on-line manner |
| CN103216311B (en) * | 2013-04-07 | 2015-01-28 | 朱晓义 | Automobile engine and power device |
| CN103306855B (en) * | 2013-06-24 | 2015-07-22 | 谷利伟 | Detonation engine |
| WO2015149167A1 (en) * | 2014-04-02 | 2015-10-08 | Fanara Roberto | Hcci rotary engine with variable compression ratio control |
| CN105547462B (en) * | 2015-12-16 | 2019-08-23 | 中国航空工业集团公司沈阳发动机设计研究所 | A kind of engine rotor radial vibration measurement method |
| CN114837802B (en) * | 2022-05-05 | 2023-10-10 | 徐继荣 | Cam rotor engine and using method thereof |
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| Publication number | Priority date | Publication date | Assignee | Title |
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| WO2018020496A1 (en) * | 2016-07-25 | 2018-02-01 | Tenenbaum Yehonatan | Split-cycle engine with a variable displacement compressor and a rotary motor |
| US10883420B2 (en) | 2016-07-25 | 2021-01-05 | Jonathan Avraham TENENBAUM | Split-cycle engine with a variable displacement compressor and a rotary motor |
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