HOMOGENEOUS CHARGE COMPRESSION IGNITION (HCCI) VANE-PISTON ROTARY ENGINE

ALLUMAGE PAR COMPRESSION D'UNE CHARGE HOMOGENE (ACCH) MOTEUR ROTATIF A PALETTE-PISTON

English 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.

French Abstract

Depuis plus d'un siècle, lorsque trois litres d'essence sont brûlés dans un véhicule équipé d'un moteur piston à allumage commandé ou d'un moteur piston diesel à allumage par compression, le 1er litre est gaspillé en chaleur, le 2e litre est perdu dans le réglage du débit au niveau du collecteur d'admission ou en énergie cinétique des gaz au niveau du tuyau d'échappement (une perte que les turbocompresseurs et le modèle de moteur à cycle thermodynamique Atkinson d'aujourd'hui ayant un taux d'expansion légèrement plus élevé tendent à réduire); seul le 3e litre est utilisé pour faire avancer le véhicule. Ce constat explique pourquoi, malgré les efforts techniques continus et les milliards de dollars dépensés en R et D, l'efficacité thermique de la plupart des moteurs piston-vilebrequin à allumage commandé et à allumage par compression stagne autour de 33 %. La hausse du prix de l'essence au cours des trois dernières décennies et les normes d'émissions atmosphériques ont poussé l'industrie automobile à étudier le mode de combustion plus prometteur qu'est l'ACCH. Le contrôle d'une combustion aussi rapide dans les moteurs conventionnels piston-vilebrequin sur une plage de vitesses utile s'est avéré plutôt difficile jusqu'ici, puisque les chambres de combustion sont conçues pour la déflagration sur le moteur à allumage commandé ou la flamme de diffusion sur le moteur à allumage par compression ou diesel, mais pas pour la détonation. Dans les moteurs à quatre temps, il faut deux révolutions de vilebrequin pour chaque temps de combustion. De plus, la faible amplitude sinusoïdale du taux de compression inhérent à la cinématique piston-manivelle à proximité du point mort haut (PMH) tend à causer soit des ratés lorsque le moteur est froid ou tourne à régime pauvre (la valeur du taux de compression est insuffisante pour la détonation à proximité du PMH), soit un cliquetis important, un excès de bruit, de vibration et de rudesse (NVH) et un taux d'accroissement de pression qui risque d'endommager la structure pendant les périodes transitoires, lorsque le moteur devient chaud ou tourne à charge plus élevée à un nombre plus faible de tours par minute (t/m) (problème de détonation prématurée). Par conséquent, la majorité des concepts d'ACCH en cours d'élaboration sont fondés sur des moteurs de type piston-vilebrequin équipés de systèmes informatisés sophistiqués qui tournent en mode ACCH en charge partielle ou au ralenti, mais reprennent, à pleine charge ou lors d'un démarrage à froid, le mode centenaire d'allumage commandé ou d'allumage par compression, lequel est moins efficace sur les plans thermique et environnemental, mais moins dur pour la structure. Cette invention vise à contrôler l'ACCH dans toutes les conditions de fonctionnement, réduisant ainsi au minimum les émissions atmosphériques du moteur et les pertes de chaleur au moyen d'un moteur palette-piston : une pompe à palettes radiales, équipée de palettes guidées positivement par des cames internes et externes, comprime graduellement l'air, ou un mélange homogène air-carburant, tout en maintenant l'air ou le mélange en deçà du niveau d'auto-allumage. Les pistons radiaux placés dans le rotor et synchronisés mécaniquement au mouvement des palettes génèrent une augmentation plus brusque du taux de compression, ce qui détonne la charge à proximité d'un angle optimal de l'arbre du rotor. Les gaz produits prennent alors rapidement de l'expansion, réduisant ainsi au minimum les pertes de chaleur au niveau des parois de la chambre pendant le temps de combustion. Ce concept offre une plage de couples utile à un nombre relativement faible de t/m, tirant avantage du principe du moteur à palettes, selon lequel il y a autant de temps de combustion que de palettes de rotor par révolution d'arbre. En raison de leur nombre élevé, les chambres de combustion peuvent avoir un volume beaucoup plus petit que dans un moteur piston-vilebrequin pour une puissance sur l'arbre similaire, ce qui est essentiel pour le mode ACCH à pleine charge. Une synchronisation mécanique simple et fiable avec les mouvements des palettes et du piston est assurée, pour un contrôle précis de détonation de la charge sur toute la plage de fonctionnement du moteur. Contrairement à la combinaison piston-vilebrequin, ce moteur palette-piston peut être conçu selon des valeurs inégales de taux de compression et de taux d'expansion : un taux d'expansion beaucoup plus élevé que le taux de compression, suivi d'une phase de recompression postallumage refroidie, maximise la puissance sur l'arbre au moyen d'un cycle thermodynamique inédit, dont l'efficacité peut dépasser celle du cycle Atkinson et très bien convenir aux applications de groupes électrogènes, de véhicules de surface et d'avions subsoniques. Un taux de compression supérieur au taux d'expansion assure une génération d'impulsions haute pression et haute fréquence (plus de cent Hertz) aux sorties d'échappement pour contrôler les moteurs à détonations pulsées pour la propulsion haute vitesse des aéronefs.

Claims

CLAIMS

The embodiments of the invention in which an exclusive property or privilege
is claimed
are defined as follows:


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.

Description

CA 02620602 2011-10-06
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.

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CA 02620602 2011-08-18
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.

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CA 02620602 2011-10-06
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.
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CA 02620602 2011-08-18
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.

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CA 02620602 2011-10-06

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).

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CA 02620602 2011-10-06

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.

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CA 02620602 2011-08-18

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).

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CA 02620602 2011-08-18

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.
-8-


CA 02620602 2011-10-06

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)-
(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.

-9-

Representative Drawing