Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
2~597~7
The present invention relates to internal co~bus~i~
engines, and in particular to a rotary internal com~us~larr
engine.
It is common in internal combustion engines to use
parts that have reciprocating and/or eccentric oscillating
movements. For example, the pistons and valves of a
conventional auto or diesel engines have a generally linear
reciprocating motion, while the rotor of Wankel-type engine
has an eccentric oscillatory motion. In such engines,
reciprocating or oscillating parts are subject to high
accelerations, and thus experience large inertial forces,
especially at higher revolution speeds. Typically, large
counterweights are required to control vibrations caused by
unbalanced inertial forces acting on the various moving
components. The counterweights and rugged construction
required for these reciprocating or oscillating parts
considerably increase the weight of the engine. Furthermore,
in spite of the use of balancing counterweights, these engines
often suffer from vibration problems, which result in losses
of efficiency.
Engines with reciprocating pistons typically require
complex mechanisms with a large number of components, thus
creating increased friction, wear, weight, production cost and
maintenance.
Four cycle engines (i.e. conventional reciprocating
piston engines) tend to be mechanically inefficient (i.e. have
a low power to weight ratio) because each piston generates
power over less than 1/2 revolution for every 2 revolutions
of the crank shaft. Furthermore, low torque is developed at
the beginning and end of the combustion cycle, due to the
minimal moment-arm about the crankshaft at these points in the
cycle.
In two cycle piston engines, the power to weight
ratio is better than in four cycle engines, but these engines
often suffer from inadequate exhausting of burned gasses and
the resultant mixing of exhaust gasses with incoming fuel/air
mixture.
2 2059757
In reciprocating piston engines, the use of intake
and exhaust valves are required to control the flow of gasses
through each cylinder. For example, the exhaust valve serves
to constrain the release of combustion gasses, with the result
that exhaust gasses must be exhausted from the cylinder within
1/2 revolution of the crankshaft.
In single stage engines (where gasses are compressed
and burned in same chamber), external super-chargers are often
employed to increase compression of the fuel/air mixture and
thereby increase the power of engine. However superchargers
consume engine power (thus reducing efficiency) and increase
the complexity of the engine.
In reciprocating piston engines and most rotary
engines, cooling of the piston (or rotor, as the case may be)
is difficult because of shape and/or movement of the piston.
In these engines, cooling is generally limited to the walls
of the combustion chambers, thus leaving the engine vulnerable
to thermal-expansion problems and overheating of internal
parts.
In order for a conventional internal combustion
piston or rotary engine to operate, highly effective seals
must be provided. These seals are critical, particularly in
the combustion chamber, where gas pressures prior to
combustion can be in the range of 10 to lS atmospheres, and
where the gas pressure after combustion can be as much as a
factor of ten~higher. However, efficient operation of a
conventional piston or rotary engine requires that the engine
match (as closely as possible) the thermodynamic "Carnot-
cycle", which demands that combustion occur within a constantvolume. It will be apparent that leakage from the combustion
chamber, particularly during and immediately following
combustion, effectively prevents constant-volume combustion,
and thus seriously affects the engine efficiency. However,
a problem common to most rotary piston engines is the
construction of effective sealing components. The seals
usually wear out rapidly, and in some cases are very difficult
.
3 20597~7
to build (for example, where inter-engaging rotors or vanes
are employed).
An object of the present invention is to provide an
internal combustion engine which avoids many of the above-
noted deficiencies.
According to an aspect of the invention, there is
provided an internal combustion engine comprising: at least
two rotary units, each of said units comprising a profiled
rotor rotatably mounted on a shaft within a cylindrical
housing, the profile of said rotor being adapted to
operatively receive therein a sector of a rotatable disc-
shaped abutment having an axis of rotation substantially
perpendicular to said rotor, which abutment includes a
substantially rectangular cutout and is rotationally coupled
to said rotor so that the rotor and the abutment rotate at the
same rate, said rotor further including a vane having a
generally twisted-helical shape whereby the vane engages into
the cutout of the abutment during each revolution of said
rotor; and a transfer duct for conducting compressed gasses
between respective ones of said rotary units.
According to a preferred embodiment of the present
invention, two rotary units are coupled together by a shaft,
and connected by a transfer duct. In this embodiment, a first
rotary unit operates as a compressor, while the second unit
operates as a combustion chamber. In addition, a rotary valve
may be used to control the passage of air through the transfer
duct.
In the following description, the rotary unit which
operates as a compressor, will be referred to as the
compressor or compression stage, while the rotary unit which
operates as a combustion and/or expansion chamber, will be
referred to as the expansion chamber.
4 20597~7
According to the present invention, all of the
primary moving parts are rotary, thus effectively eliminating
component stress and vibration due to non-circular movements.
As a result, the weight of the engine can be reduced because
the reduction of inertial stresses, due to the elimination of
reciprocating and/or eccentric motions, permits lighter
construction of the supporting structures of the engine, and
limited use of counterweights.
The simplicity of a true rotary engine reduces
mechanical complexity and the number of moving parts thus
reducing friction, wear, weight, cost of production and
maintenance.
The engine according to the present invention is
capable of generating power over more than 2/3 of each
revolution of the shaft. Additionally, maximum torque is
obtained throughout the combustion/expansion phase because the
moment arm of the expansion thrust (against the face of the
expansion chamber vane) remains constant throughout the
expansion cycle. In effect, the advantages of the two cycle
and four cycle engines are combined in the present invention,
as exhaust gasses are not mixed with intake gasses within the
combustion chamber, and the engine generates power during each
revolution of the shaft.
Within the engine of the present invention, fresh
air is drawn into the compressor stage; compressed to elevate
its pressure and temperature; and then transferred to an
expansion phase. During the transfer and expansion processes,
fuel is injected into the air flow, ignited, and the burning
air/fuel mixture allowed to expand at substantially constant
pressure. In this respect, the engine of the present
invention operates approximately according to the "Joule
~ycle", which is well known to be the thermodynamic cycle by
which conventional gas-turbine engines operate. The present
invention retains much of the simplicity and mechanical
efficiency of the gas-turbine engine (in terms of the small
number of moving parts) but has significant advantages over
gas turbine technology. In particular, the engine according
20~97~7
to the present invention is capable of maintaining
substantially higher compression ratios than is possible with
conventional gas turbines, thus the operating efficiency is
similarly much higher. Additionally, the engine of the
present invention is capable of operation with substantially
less air flow, and at substantially lower operating speeds,
than is required for successful operation of a gas turbine.
The engine according to the present invention does
not require an exhaust valve, because the interaction between
the vane and abutment of the expansion chamber provides a
physical separation between the gasses currently expanding
and/or burning within the expansion stage, and exhaust gasses
produced during the previous rotation cycle of the engine.
Thus the exhaust gasses can be expelled continuously, without
being constrained by the use of an exhaust valve. The
elimination of exhaust valves reduces the number of moving
parts.
The use of separate compression and expansion stages
permits the use of a compressor stage having (for example) a
larger volume than the combustion stage, thus facilitating
increased compression ratios in the combustion/expansion
stage.
The rotors, vanes, abutments and housings can be
continuously cooled and maintained at a similar temperature,
thus preventing the overheating of internal parts.
Maintaining all parts at similar temperatures reduces stress
and adjustment problems caused by variations in the thermal
expansion of adjoining parts.
30The present invention will now be described, by way
of example, with reference to the appended drawings, in which:
Figure 1 shows an embodiment of a rotor and an
abutment according to the invention;
Figure 2 shows an embodiment of an engine according
35to the present invention, illustrated in a first position in
its operating cycle.
6 20597~7
Figure 3 shows the embodiment of Figure 2 at a
second position in its operating cycle;
Figure 4 shows the embodiment of figure 2 at a third
position in its operating cycle;
5Figure 5 shows the embodiment of Figure 2 at a
fourth position in its operating cycle;
Figure 6 illustrates the exterior features of the
housing of the embodiment of Figure 2;
Figure 7 presents an illustration of the curve
10mapped out by the rotation of an abutment in association with
a rotating rotor;
Figure 8 illustrates the alignment between the face
of an abutment and the centre axis of a rotary unit employed
as a compressor;
15Figure 9 presents a cross-sectional view taken
through a rotary unit according to the invention;
Figure 10 illustrates an abutment housed within an
abutment housing;
Figure 11 presents a partial cross-section view of
20a vane and a rotary unit housing, illustrating sealing and
lubrication arrangements;
Figure 12 presents an illustration of an embodiment
of a transfer duct according to the present invention;
Figure 13 presents a side view of the transfer duct
25illustrated in Figure 12;
Figure 14 presents a diagrammatic cross-sectional
view through a rotary unit according to the present invention,
employed as an expansion and combustion chamber;
Figure 15 presents an illustration of the rotor and
30vane of the rotary unit of Figure 14;
Figure 16 illustrates an abutment and an abutment
housing employed in the rotary unit illustrated in Figure 14;
Figures 17A - 17D illustrate various embodiments of
mechanical seals employed to provide sealing between a rotor,
35abutment and vane in an embodiment of the present invention;
7 20~9757
Figures 18A and 18B illustrate the relationship
between an abutment edge and adjacent rotor and vane surfaces
in an alternate embodiment of the present invention.
Figure 19 illustrates an embodiment of shaft and
gear arrangements used to provide synchronised rotation of the
various components in an embodiment of the present invention.
Figure 20 presents an illustration of the effect of
increasing the size of a rotor; and
Figure 21 presents an illustration of the effect of
increasing the size of an abutment.
It should be noted that the following discussion
of example embodiments of the invention includes discussions
of possible mechanical sealing, lubrication and cooling
arrangements which, in a practical engine, would be required
to ensure smooth and reliable operation. However, it will be
understood that the precise design details of these
arrangements are considered to be in the domain of those
skilled in the art. Accordingly, it will be understood that
the possible mechanical sealing, lubrication and cooling
arrangements, described in conjunction with the example
embodiments, are not to be taken as being limitative of the
present invention.
According to an embodiment of the present invention,
the compression and expansion rotors are coupled together and
therefore rotate simultaneously (see Figures 2 to 5).
However, to simplify the description of the compression,
transfer, expansion and exhaust phases of operation, each
phase is explained separately.
Figure 1 shows a rotary unit and an associated
abutment. In this case, the rotary unit operates as a
compressor, and the rotor is illustrated at an intermediate
(relatively early) point during the intake/compression cycle.
As the compression rotor 14 rotates (in the direction of the
arrow), fresh air is drawn into the expanding intake volume
2 of the compression stage, through an intake port 1. At the
same time, air which entered the compression stage during the
~0~97~7
previous rotation of the compression rotor 14 is compressed
between the advancing leading face 47 of the compressor vane
3 and the high-pressure side 4h (see Figure 2) of the
compressor abutment 4.
Figure 2 shows a compressor, transfer duct and
expansion chamber at a point near the end of a compression
cycle. A rotary transfer valve 6, which is rotationally
synchronized with the compressor rotor 14, seals the transfer
duct inlet 7 between the compression stage and the transfer
duct 5.
Referring to Figure 2, at the end of the compression
cycle the compressor rotor 14 and abutment 4, which are also
rotationally synchronized, are in a position where the
compressor vane 3 is about to engage into a substantially
rectangular compression abutment gap 4g. The rotary valve 6
is also in a position where the valve 6 is about to open the
transfer duct inlet 7.
At this point, the combustion vane 9 almost
completely blocks the passage between the transfer duct 5 and
the expansion chamber 8, which is (in this position) confined
to a very small volume defined by the expansion vane 9 and the
high pressure side lOh of the expansion abutment 10, which
engaged with each other at a slightly earlier point in the
rotation of the engine.
As illustrated in figures 2 to 5, the compression
stage can have a larger volume than the expansion chamber,
thereby allowing for further compression of the gasses during
transfer, and to compensate for the volume of the transfer
duct 5.
The compression ratio for the engine can be
calculated according to:
C = Vl - tV2 + V3?
V4 + V5 + V6
Where: C = Compression ratio
V1 = Volume of compression chamber
V2 = Volume of vane
V3 = Volume loss due to intake port
20~9757
V4 = Volume remaining in compressor at the end of
transfer cycle (as explained below).
V5 = Volume of transfer duct
V6 = Volume of expansion chamber at end of transfer
cycle (explained below)
The volumes Vl - V6 can be calculated from
geometrical considerations. By using the above relationship,
and assuming that compression loss due to leakage is
negligible, the compression ratio of the example engine
illustrated in Figures 2 through 6 can be determined to be
approximately 7.5:1. It will be understood that under true
operating conditions, a certain degree of "blow-by" (or gas
leakage around the abutments and vanes) is inevitable. In
an embodiment of the present invention, the amount of blow-
by is restricted by the use of sealing elements adapted to
provide a mechanical seal between moving parts.
In an alternate embodiment, the blow-by is limited
by the shape of the edges of the abutments and vanes and by
running the engine at sufficient speeds to limit the time
available for air leakage so as to reduce the compression
losses. In this alternate embodiment, a narrow running
clearance is maintained between the various components. Thus
while the amount of blow-by is substantially higher than in
a mechanically sealed embodiment, the internal friction and
wear of the engine is greatly reduced. Additionally, the
mechanical simplicity of the engin is enhanced by the
elimination of mechanical sealing elements. Maintenance of
an operating clearance between components also has the
advantage of obviating the problems associated with
lubricating vane and abutment surfaces, because frictional
contact between these surfaces is eliminated.
Figure 3 illustrates the engine illustrated in
Figure 2 at an intermediate point during the transfer phase
of operation. At this point, the rotary valve 6 is open, thus
allowing compressed air (or fuel-air mixture) to pass through
the transfer duct 5 and into the enlarging expansion chamber
~3 defined by the expansion abutment 10 and expansion vane 9.
lO 20~9757
Note that Figure 3 also shows a different
arrangement of the structural connection 38 between the
compression and expansion stage housings.
Figure 4 illustrates the engine illustrated in
Figure 2 at the end of the transfer phase of operation. At
this point, rotation of the rotary valve 6 has closed the
transfer duct inlet 7. The compressed air (or fuel-air
mixture) fills the transfer duct 5 and the continuously
enlarging expansion chamber 8 defined between the expansion
vane 9 and the expansion abutment 10.
Fuel can be mixed with the air either prior to
entering the compressor, or injected in the transfer duct 5
during the transfer cycle. A fuel/air mixture can enter in
the compressor via a conventional carburettor, which
facilitates improved fuel vaporization and mixing of gasses
before entering the transfer duct and expansion chamber. In
this case it is also advantageous to use either fuel which in
a gaseous form at operating temperatures and pressures (for
example, propane, natural gas etc.), to avoid dissolving of
lubricant used in compression chamber parts. Alternatively,
two cycle oil can be added to the fuel to serve as lubricant,
or the liquid fuel could be heated to promote vaporisation.
Otherwise a gasoline resistant lubricant should be used to
lubricate compression stage components.
Alternately, fuel (liquid or gaseous) can be
injected into the transfer duct 5 during the transfer phase.
In this case, the injector should be placed as close as
possible to the transfer duct inlet 7 to allow more time for
mixing of the fuel and air, prior to entering the expansion
stage.
At the end of the transfer cycle, a spark plug 11,
located in the transfer duct 5, ignites the fuel-air mixture
so as to initiate expansion. A second spark plug 59 (see
Figure 6) can be installed within the expansion chamber wall
(a suitable recess can be provided to avoid obstructing the
passage of the expansion vane), close to the expansion
11 2059757
abutment 10, to facilitate direct ignition of the fuel within
the expansion chamber.
Ignited gasses, which are sealed within the
expansion chamber 8 by the expansion abutment 10, the rotary
valve 6 and the expansion vane 9, exert a high and
substantially constant, pressure on the trailing face 48 of
the expansion vane 9, while the leading face 49 of the
expansion vane 9 is subject to much lower (approximately
atmospheric) pressure. The pressure difference across the
expansion vane 9 produces a torque about the expansion rotor
26 and output shaft 13, thereby forcing the expansion rotor
26 to rotate in the direction indicated by the arrow in Figure
4.
Thrust generated on the expansion vane 9 decays
towards the mid-point of the next transfer cycle (see Figure
3) as the gasses in the expansion chamber expand to near
atmospheric pressure, and both leading 49 and trailing 48
faces of the expansion vane 9 are subject to approximately the
same pressure. Thus thrust is generated on the expansion vane
9 for approximately 235 with a decrease in thrust for the
following 60, as the expansion vane 9 progressively engages
with the expansion abutment 10. Power is therefore generated
over more than 2/3 of each revolution.
An exhaust port 12, is located in the lower part of
the expansion chamber, close to the expansion abutment 10, as
shown in Figures 5 and 6. The exhaust port 12 does not have
a valve, and thus burned gasses are expelled continuously
through the exhaust port 12 as the leading face of the
expansion vane 9 advances towards the port 12.
Figure 6 illustrates the exterior of the compressor
and expansion chamber housings, clearly indicating the
relative positions of the compressor inlet 1, transfer duct
inlet 7 and outlet 25, and the exhaust port 12. In addition,
the slots 44 and 45, through which the compressor and
expansion abutments respectively are mounted, are illustrated.
Furthermore, the position of the rotary valve housing 43, with
respect to the transfer duct inlet 7 and the compressor
20597~7
12
abutment is illustrated. Finally, an opening 59, can be
provided to facilitate installation of a spark plug so as to
permit ignition of fuel/air mixture within the expansion
chamber, as discussed previously.
The offset between the compression and expansion
components (i.e. the angle between the compression and
expansion abutments) is dictated by the length of the transfer
cycle and the synchronization of respective compression and
expansion rotors and vanes. In the embodiment shown in
Figures 2 through 6, the transfer cycle begins before the
expansion vane has completely crossed the outlet 25 of the
transfer duct 5 (to enclose the expansion chamber), so as to
permit the incoming compressed gasses to force some of the
exhaust gasses within the transfer duct to be vented past the
expansion vane, and thus reducing the amount of exhaust gasses
being introduced into the expansion chamber 8 from the
transfer duct 5.
The accurate alignment of the abutments with the
respective rotors is of prime importance, because the
alignment determines the degree to which an adequate seal can
be maintained to prevent excessive blow-by of gasses around
the abutments, with the associated losses in compression and
efficiency and utilisation of expansion chamber pressure.
Figure 7 illustrates the curvature of the faces of
the compression and expansion vanes. The rotation of the
rotor and abutment are synchronised, so that the contour of
the vane must correspond with the motion of an edge of the
abutment gap as the rotor and abutment rotate together. As
illustrated in Figure 7, a point (a) at a given radius from
the centre of the abutment (a point at the perimeter of the
abutment is illustrated) maps out a curve F as the rotor and
abutment rotate in the directions indicated by respective
arrows. Locations b, c, d and e on the curve F indicate
various positions of the point (a) at various stages during
the rotation, and illustrate the relationship between the
plane of the abutment face G and the vane face at those
locations. In particular, it can be seen that the angle
13 ~059757
between the abutment face G and the vane varies as the vane
passes through the abutment.
Because the angle between the abutment face G and vane
face F varies along the path followed by the two components,
it is not possible for both leading and trailing faces of a
vane to conform with the leading or trailing faces of an
abutment.
The alignment between the abutments and rotors in
the engine according to the invention is such that a seal must
be maintained between one edge on the high-pressure side of
an abutment and one face of the corresponding vane. In
particular, the compressor abutment 4 must maintain a seal
against the leading face of the compressor vane 3, while the
expansion abutment 10 must maintain a seal against the
trailing face of the expansion vane 9. Referring to Figure
8, the high pressure side 4h of the compression abutment 4,
which faces the leading face of the compression vane 3, is
aligned with the centre axis of the compression rotor 14.
Conversely, the side of the expansion abutment which faces the
trailing face of the expansion vane 9, is aligned with the
centre axis of the expansion rotor.
Figure 9 shows a cross-section of the compression
rotor 14, the compression stage wall 15 and rotor sealing
rings 16. The rotor can be formed as a hollow shell,
contoured 17 to correspond to the shape of the abutment.
Counter-weights 18 can be placed inside the rotor, opposite
to the compressor vane 4, to balance the weight of the vane,
thus eliminating unbalanced forces, and associated vibrations.
The rotor rings-seals 16 are stationary (i.e. do not rotate
with the rotor), a portion of each ring seal 16 being engaged
into the chamber housing 15. The rotor seals 16 can be made
of steel to allow for flexibility and wear resistance
properties. The ring-seal gap 53 is located in line with the
abutment 4 to prevent any leakage of gasses and permitting
lubrication of both rotor seal and abutment seal at the same
location.
14 20597~7
Referring to Figure 10, the gap 4g in the compressor
abutment 4 is shaped to accept the compressor vane 3 while
passing through the abutment 4, with only one edge 50 (on the
high-pressure side 4h) closely fitting the face of the vane.
Since the curves along inside and outside edges of the vane
3 are different, the inside faces 21 of the abutment gap 4g
can be twisted in order to more closely conform to the face
of the vane. However, because only one edge 50 must maintain
a seal with a corresponding face of the vane 3, the precise
geometry of the remaining surfaces of the abutment gap is not
critical, beyond the fact that they must not interfere with
passage of the vane 3.
A seal, made of compressible resistant material,
is disposed near the sealing edge 50 of the gap 4g and along
the circumferential edge 22 of the abutment 4, in order to
seal the gaps between the abutment, rotor and vane. The seals
need to be compressible to provide effective sealing using a
single seal between the abutment and the vane, as only the
high-pressure face 4h of the abutment 4 is aligned with the
centre axis of the rotor, and only one edge 50 of the abutment
4 exactly follows the curvature of the face of the vane 3.
Using a compressible material permits a wider area of seal
contact, and allows for compensation of thermal expansion of
parts.
A preferred material for abutment seals would be a
silicone-teflon base substance for it's heat resistance and
lubrication properties. The material should however be hard
enough to resist pressures exerted by the compressed gasses
and needs to be only slightly compressible as the gaps to be
filled due to thermal expansion are extremely small
(hundredths of millimetre).
A housing 23 is placed around the abutment to keep
pressure on the seal while not in contact with the rotor to
prevent excessive deformation of the flexible seal at high
revolution speeds. The faces of the abutments are lubricated
by using oil-grooves inside the housing walls. The gap faces
15 205975~
are lubricated by injecting oil (or lubricant) through small
holes located on the inside face of the abutment gap 57 as the
gap enters the housing. Centrifugal force will cause the oil
to flow along the entire face of the seal.
The abutment housing 23 is attached to the wall of
the compression chamber and the abutment 4 enters the chamber
through a slot 44 (see figure 6) in the chamber wall. A seal
56 (see Figure 8) is placed between the abutment housing and
the compression face 4h of the abutment to prevent leakage
through the abutment gap while it is inside the housing.
The abutment can be balanced by creating hollow
areas inside the abutment opposite the gap, or by using
external counterweights attached to the abutment shaft. These
abutment counterweights would have to be located outside the
abutment housing as the housing needs to closely fit the
abutment to permit proper heat transfer for cooling purposes.
The vane 3 should be as thin as possible while
remaining strong enough to sustain pressures exerted by
compressed gasses and resist centrifugal forces caused by the
high revolution speeds of the rotor.
A wear resistant rigid seal 19 ( for example a metal
alloy~ is placed between the vane and the compressor wall 15
(see Figure 11) and the expansion gap is filled with
compressible material 20 (similar to the abutment seals) to
prevent escape of compressed gasses past the face of the
abutment during compression.
An oil groove 52, filled with porous rigid material
is located in the wall of the compression chamber 15 to oil
the vane seal 19 as it passes over the groove 52. The oil
groove is located close to the compressor inlet port 1 to
avoid being submitted to high pressures which would impair
the lubrication of the vane seal 19.
The rotor, vane, abutment, chamber wall and housing
can be made of metal having similar thermal expansion
coefficients to avoid problems related to differences in
thermal expansion. Iron and steel alloys are preferred for
this embodiment because of their low thermal expansion
16 2~5975~
coefficient, good resistance and rigidity, ease of casting
(important mainly for the rotor and vane) and low production
cost. On the other hand, aluminium alloys (such as silicon
and copper) are better heat conductors, are lighter and can
attain resistance approaching that of steel but have higher
thermal expansion coefficients and are more expensive.
Referring now to Figure 12, the rotary valve 6 is
constructed as large as possible to permit the valve gap 24
to be in the full open position as long as possible in order
to permit maximum free flow of gasses between the compressor
and expansion chamber through the transfer duct inlet 7.
A valve seal between the valve 6 and the compressor,
is not required as the valve rests tightly against the
compression chamber face of the valve housing 43 (see Figure
6) and the valve overlaps the transfer duct inlet 7. The
valve 6 is therefore forced to bear against the compressor
side of the housing 43 by the pressure exerted by the
expansion gasses, except (possibly) at the end of the
compression cycle where pressure on both sides could become
similar. As a security measure, a spring loaded washer could
be placed in the valve housing on the transfer duct side to
force the valve 6 to continuously bear against the compressor
side of the housing 43.
A spark plug 11 is placed in the transfer duct 5 as
close as possible to the exit port 25 in order to ignite the
fuel-air mixture efficiently. Figure 13 provides a side view
of the transfer duct.
The rotary valve 6 can be balanced by using methods
similar to that described in connection with the compression
abutment. Lubrication of the valve 6 can be accomplished by
providing oil grooves on the valve housing. A seal is also
placed inside the valve housing, around the entrance port 7,
on the side adjacent to the transfer duct, to prevent
expansion gasses from exerting a back pressure on lubricating
oil. This seal has to be larger than the valve gap so the
17 2~59757
seal is always resting on a portion the valve large enough to
avoid any interference between the valve and the seal.
Referring to Figure 14 and 15, the shape of the
expansion rotor 26, is similar to that of the compression
rotor, but the expansion rotor is smaller and is of heavier
construction to resist internal expansion chamber temperature
and pressure. Sealing rings can be placed at each end of the
rotor 27, the arrangement and lubrication thereof being
similar to the compressor rotor seals described previously.
Similarly, counterweights can be located within the expansion
rotor in a manner similar to that described above, to balance
the expansion vane. However, care must be taken to ensure
that the counterweights do not interfere with cooling of the
rotor.
The expansion vane 9 is shaped as described above
in connection with the compression vane 3, but is generally
thicker, and is provided with internal cooling channels to
permit effective cooling of the expansion vane.
Referring to Figure 15, a seal 28 is fitted in a
groove along the peripheral edge of the vane 9 between the
vane 9 and the expansion chamber wall, to compensate for
possible difference in expansion between the vane 9 and
chamber wall, while maintaining an effective seal.
Referring to Figure 16. the expansion abutment 10
has the same general shape as the compression abutment 4, but
has wider gap lOg to accommodate the wider expansion vane.
The high pressure face lOh is aligned with the centre axis of
the rotor, and a sealing edge 51 closely follows the trailing
face of the expansion vane 9 in order to maintain an efficient
seal in the expansion chamber 8.
The seals on sealing edge 51 of the abutment gap
lOg, and the seals around the perimeter of the abutment 10 are
preferably made of compressible heat and wear resistant
material similar to that described for the compression
components. The seals could also be made of hard and flexibly
mounted (i.e. spring loaded) material. Figures 17(a) and
20597~7
18
17(b) illustrate the use of compressible seals in the
compressor and expansion chamber abutment gaps 4g and lOg
respectively. Rigid seals, as illustrated in Figures 17(c)
and 17(d), can also be used. In the latter case, the seal
movements can be limited by spring-loaded chokers which
facilitate continuous adjustment of the sealing element, and
thus ensuring an effective seal between a vane and abutment.
A housing 33 is placed around the abutment to keep
pressure on the seal while not in contact with the rotor to
either prevent excessive deformation of the compressible seal
at high revolution speeds or keep a tight fit on the spring
loaded seal. A lining 46 can be used within the housing 33
(see Figure 14), to maintain alignment between the seal and
the rotor. A seal is placed in the abutment housing 33,
between the abutment housing 33 and the high pressure face lOh
of the abutment 10 to prevent leakage through the abutment gap
lOg while inside the housing. The abutment housing 33 is
attached to the wall of the expansion chamber and the abutment
10 enters the chamber through a slot 45 (see Figure 6) in the
expansion chamber wall.
The abutment 10 can be counterbalanced and all seals
are lubricated in the same manner as described above with
respect to the compression stage.
The expansion rotor, vane, abutment and cylinder
wall, as well as the rotary valve and transfer chamber are
preferably cooled to keep thermal expansion to a minimum thus
avoiding distortion and seizing due to differences in thermal
expansion rates between various adjoining parts. Figure 19
illustrates a centrifugal fan 34, directly coupled to the
compressor rotor, which can be used to cause a circulation of
air through and around the engin. The precise arrangements
for cooling, either by air-cooling, a circulating cooling
liquid (such as water, or at least predominantly water), or
a combination of these, is within the domain of one skilled
in the art, and is therefore not described in detail here.
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20~97~7
19
The above described sealing, lubrication and cooling
arrangements are generally related to a mechanically sealed
embodiment of the invention. As previously mentioned, in an
alternate embodiment, mechanical seals between the rotor,
vane, and abutment are not employed. In this embodiment, gas
flow through the narrow gaps between these elements is
restricted by the shape of the edge of the abutment and vane,
and by maintaining a narrow clearance between adjacent
surfaces. In particular, Figures 18A and 18B diagrammatically
illustrates the relationship between the edge of the abutment
and the rotor (or vane) surface and between the edge of the
vane and the interior surface of the rotor housing. The
geometry illustrated in Figures 18A and 18B utilises the
hydrodynamic principle that a sharp-edged inlet to a duct
increases the resistance to fluid flow through the duct. Thus
the high pressure side of the abutment (and vane) is provided
with a sharp, chisel-like, edge. The peripheral surface of
the abutment (or vane) may be parallel to the adjacent
surface, or may be angled slightly so that the clearance
increases towards the low-pressure side of the abutment (or
vane), as illustrated. In the latter case, the peripheral
surface should be close to parallel in order to provide
sufficient material to ensure adequate strength at the edge,
and to facilitate adequate conduction of heat away from the
edge. In addition, a narrow operating (or running) clearance
is maintained to prevent frictional contact between the edge
and the adjacent surface. This clearance should be as narrow
as possible, while being large enough to ensure that a minimum
clearance will be maintained in spite of thermal expansion.
In the alternate embodiment, a certain amount of
compression loss is inevitable. However, the adverse effects
of the compression loss can be at least partially compensated
for by increasing the rotation speed of the engine, thereby
reducing the amount of time available for compression loss
during each rotation cycle of the engine. Additionally, as
the engine heats up, thermal expansion of the rotors, vanes,
abutments and housings reduces the width of the running
'~3,
2~97~7
clearance. Thus the amount of compression loss reduces as the
engine is brought up to normal operating temperature.
Elimination of mechanical seals and maintenance of
a working clearance, substantially simplifies the engine, and
reduces friction, thereby obviating the need for lubrication
of the edges of the abutments and vanes. However, the higher
operating speeds, and the fact that the engine is now running
"dry" (i.e. without lubrication) can be expected to increase
the operating temperatures of the engine. In addition,
because the amount of compression loss is directly related to
the clearance between components, which is in turn dependent
on thermal expansion rates, careful selection of heat-tolerant
materials is very important. For example, the abutments,
rotors and vanes can be fabricated from cast ceramics of the
type currently being used in advanced gas-turbine engines, and
(experimentally) in automotive piston engines. These
materials exhibit low thermal expansion rates, high resistance
to thermal shock, and have also been shown to maintain
dimensional tolerances and structural integrity at extreme
temperatures.
Finally, the superior thermal resistance of cast
ceramics means that the abutments, and rotor vanes can be
operated substantially without cooling. Thus in this
alternative embodiment, cooling of the engine can essentially
be limited to the circulation of air around the exterior of
the engine (possibly using cooling fins to increase heat
transfer rates), and (possibly) the interior of the rotors.
Alternatively, liquid cooling can be used, in this case by
circulating cooling liquid (such as, for example, a
conventional coolant consisting at least predominantly of
water) through a jacket surrounding the transfer duct and
expansion chamber.
As shown in Figures 2 through 5, the compression 14
and expansion 26 rotors are directly linked together. The
expansion 10 and compression 4 abutments and rotary valve 6
are linked by suitable gears and shafts to the rotors 14 and
2059757
21
26, so that all moving parts are synchronized. For example,
the shaft and gear arrangements shown in Figure 19 illustrates
an embodiment of the invention which includes a single
synchronising shaft, and appropriate gears to ensure that the
compressor and expansion chamber abutments 4 and 10, as well
as the rotary valve 6, all rotate together at the correct
rotation speeds by being driven by a gear coupled to the
output shaft 13 of the engine. The precise design of the
gears, shafts, bearing arrangements, lubrication and gear
housings etc. is considered to be well within the capabilities
of those skilled in the art, and therefore is not discussed
in further detail here.
The rotors are mounted within the housings using
suitable bearings. Seals (for example ring seals) are
provided at either end of the rotors to prevent escape of
gasses around either end of the rotor. Here again, the
precise design of the rotor bearings and end-seals, including
supporting structures, housings, lubrication etc. is well
within the capabilities of those skilled in the art, and
therefore is not discussed in detail here.
It will be apparent from the forgoing that there
will be a variety of ways in which the engine of the invention
can be modified. The following is a brief description of
some of the modifications which are possible.
As shown in Figure 20, the size of the compression
chamber can be modified by increasing or decreasing the axial
length of the rotor 14 without increasing the diameter of the
abutment 4, and thereby without affecting the alignment of
compression and expansion abutments and their associated
gearing system.
The shape of expansion and compression chambers
shown in Figures 2 to 5, represents a preferred embodiment of
the invention. It is possible to use other shapes of vane and
chamber or proportions between vane, rotor and abutment.
However, variations in these components will influence the
.~.
20597~7
22
performance and characteristics of the engine as explained
below.
Figure 21 shows the effect of increasing the
diameter of the abutment from A to A1. Since the rotation of
the abutment is synchronized with the rotor ~and vane) B, the
length of the transfer cycle is proportional to the angle n,nl
formed by the contact between the rotor and the abutment.
Thus increasing the diameter of the abutment decreases the
circumferential length of the transfer cycle without
increasing significantly the volume of the chamber. Figure
18, shows an increase of the abutment diameter of 20%, which
produces a decrease in the circumferential length of the
transfer cycle by 11% (note that n>nl) and an increase in the
volume of the compression chamber by only 5%. A shorter
transfer cycle would increase internal pressure in the
compression chamber resulting in excessive pressures on the
seals thus increasing the risks of loss of compressed gasses.
Conversely, increasing the size of the rotor (as
shown in Figure 20), has the effect of increasing the
circumferential length of the transfer cycle and increasing
the volume of the chamber. Increasing the size of the
compression rotor would increase the compression ratio, but
an increase of the expansion rotor could lead to problems
related to thermal expansion of the abutment inside the
cylinder and increase the difficulty of sealing the gap
between the abutment and rotor. It would also create problems
related to a shorter cooling cycle of the abutment and would
require a stronger abutment due to the increased area
subjected to expansion pressure.
An increase in the diameter of both compression and
expansion rotors would increase the volume of the chambers,
the circumferential length of the transfer cycle and increases
the torque of the engine. On the other hand an increase in
rotor diameter increases centrifugal force on the vane seal,
the length of rotor seal and seal area of abutment face on the
vane. These potential problems could, however, be compensated
for by doubling the rotor rings and counterbalancing the
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,
23 2~597~7
vane seal.
Reducing the rotor to the shape of a disk, as
described in United States Patent Nos, 3,841,276, 3,502,053,
3,674,982, 3,012,551 and 3,205,874 has the disadvantage of
creating a chamber volume for a given diameter of rotor and
makes the sealing of the rotor extremely difficult,
particularly where the abutment enters the rotor housing.
The compression and expansion stage walls could also
be convex, thus providing each stage with a barrel shaped
exterior appearance. The advantage of using convex chamber
walls is the increase chamber volume with little increase in
engine size. The disadvantages are related to increased
problems of thermal expansion of a large expansion vane inside
the chamber; the complexity of vane seal fitting with the
rotor seal; the necessity of building the stage housings in
two sections facilitate insertion of the rotor; and increased
problems in cooling the vane due to it's complex shape.
It is also possible to use two or more abutments
arranged around a rotor. However, this would yield little
advantage as the total engine displacement would be the same,
and the engine would go through two or more "dead" phases
(corresponding to the transfer cycle), instead of only one.
In order to increase engine output while maintaining
continuous power (i.e. minimising any dead phase), the
preferred embodiment uses a second set of compression and
expansion stages, connected in line with the first set, and
offset by 180~ about the common shaft.
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