ROTARY PISTON STEAM ENGINE WITH BALANCED ROTARY VARIABLE INLET-CUT- OFF VALVE AND SECONDARY EXPANSION WITHOUT BACK-PRESSURE ON PRIMARY EXPANSION

MACHINE A VAPEUR A PISTON ROTATIF AVEC SOUPAPE D'ISOLEMENT D'ADMISSION VARIABLE ROTATIVE EQUILIBREE ET DETENTE SECONDAIRE SANS CONTRE-PRESSION SUR DETENTE PRIMAIRE

English Abstract

Rotary piston steam engine with equal double rotary pistons is provided with a
balanced rotary variable inlet
cut--off valve for enhanced efficiency. The exhaust steam from the primary
expansion is routed to secondary expansion avoiding back
pressure for additional efficiency. The rotary valve has balanced dual inputs
and outputs on opposite sides. The exhaust steam
from the primary expansion is taken off when the trailing face of the rotary
piston passes the inlet port of the expansion chamber
housing, the exhaust outlet secondary expansion being placed approximately 180
degrees from the primary expansion inlet in the
curved portion of the expansion chamber housing wherein back pressure is not
imparted to the primary expansion.

French Abstract

La présente invention concerne une machine à vapeur à piston rotatif ayant des doubles pistons rotatifs égaux et une soupape d'isolement d'admission variable rotative équilibrée pour améliorer le rendement. La vapeur d'échappement de détente primaire est dirigée vers une détente secondaire, évitant une contre-pression et améliorant encore le rendement. La soupape rotative comprend des entrées et sorties doubles équilibrées sur des côtés opposés. La vapeur d'échappement de détente primaire est prélevée lorsque la face de fuite du piston rotatif passe l'orifice d'admission du carter de chambre de détente, la sortie d'échappement de détente secondaire étant placée approximativement à 180 degrés de l'admission de détente primaire dans la partie incurvée du carter de chambre de détente où la contre-pression n'est pas communiquée à la détente primaire.

Claims

Claims.

1. We claim to have invented within the field of mechanical engineering,
improvements
in the energy efficiency of steam powered equal double rotary piston power
plants,
whereby more full use of the expansion of steam, or other suitable
compressible
working fluid, within a close fitting expansion chamber, is effected;

a, firstly, by using a balanced rotary variable inlet cut-off valve with
balanced dual inputs
and outputs on opposite sides of the valve,

b, secondly, by using exhaust steam from the primary expansion of the equal
double
rotary piston power plant being taken off when the trailing face of the rotary
piston
passes the inlet port of the expansion chamber housing, the exhaust outlet
secondary
expansion being placed approximately 180 degrees from the primary expansion
inlet in
the curved portion of the expansion chamber housing, wherein back-pressure is
not
imparted to the primary expansion, and

c, thirdly, steam leakage between the flat surfaces of the rotary pistons and
the adjacent
surface of the expansion chamber housing is prevented by a particular
arrangements of
seals and grooves,

d, comprising combinations of at least one of the following claims:

2. The balanced rotary variable inlet cut-off valve of claim 1, wherein said
valve includes
a cylinder rotating within a housing having two pairs of inlet and outlet
ports, the cylinder
having a plurality of pairs of grooves formed circumferentially around the
cylinder, the
plurality of grooves corresponding to the predetermined number of inlet cut-
off settings
that are to be used in a particular application, and the pattern of inlet and
outlets
alternating around the circumference, wherein the classic water wheel or
turbine like
effect of steam entering the valve assists rotation of the cylinder in a
constant direction.
3. The grooves of claim 2 are orientated in planes normal to the axis or
rotation of the
cylinder of claim 2, and the grooves extend a predetermined fraction of 180
degrees
around the cylinder, the predetermined fraction being the same fraction as
that desired
for inlet cut-off of steam, an example being 50 percent inlet cut-off having
two grooves
in the same plane, each extending 90 degrees, and spaced evenly around the
circumference of the cylinder of the rotary valve, furthermore a single groove
may be
formed extending a full rotation around the cylinder of the rotary valve,
wherein full
steam pressure in applied continually to the expansion chamber.

4. The grooves of claim 2 and the leading edge of the recesses of claim 7, are
aligned
so that the portion corresponding to the start of inlet cut-off are aligned
substantially in a
straight line.


18



5. The pairs of grooves of claim 2, and the non circumferential edges of the
recess of
claim 7, each have a shape that is aerodynamically curved to minimize
turbulence in
the high velocity high pressure steam on entry into, transit through, and exit
out of the
valve, the shape of the curve in a cross sectional plane,

a, normal to the rotational axis of the cylinder, being two short curves of
relatively small
radii of curvature, not necessarily of the same curvature, meeting with one
longer chord-
like curve of relatively larger radius of curvature, all curves generally
being circular arcs
for simplicity of manufacture, but without excluding other suitable
aerodynamic
contours, the two short curves meeting with the surface of the cylinder at an
angle
substantially in line with the holes forming inlet and outlet ports as
described later in
claim 8, the angles of inlet and outlet generally, but not necessarily being
equivalent, to
the angle of the inlet and outlet ports as they pass through the housing as in
claim 8,
b, the shape of the grooves of claim 2 or recesses of claim 7, in the cross
section of a
plane that includes the rotational axis of the cylinder, having the sides of
the groove or
recess exit the surface at an angle that is substantially normal to the
surface of the
cylinder, and having the base of the groove or recess connected to the side
walls,
preferentially in a aerodynamically smooth contour.

6. The number of distinct grooves of claim 3 have a plurality corresponding to
the
number of predetermined settings of inlet cut-off, for example, 100 percent,
50 percent,
20 percent and 10 percent in a four setting inlet cut-off valve, the pairs of
such grooves
being distributed evenly along the axis of rotation of the of the cylinder
with
approximately equal spacing between the grooves, allowing between each pair of

grooves a suitable thickness of material for containing steam under pressure,
and with a
shoulder at each end of the cylinder broad enough to ensure stability of the
cylinder on
high-speed rotation inside the valve housing described in claim 9, whereby
wear is
evenly distributed and thus reduced.

7. The balanced rotary variable inlet cut-off valve of claim 1, wherein said
valve includes
a cylinder rotating within a housing having two pairs of inlet and outlet
ports, the cylinder
having a single pair of equal three-sided recesses, rather than plurality of
pairs of
grooves, the recesses being formed on the cylinder's outer surface, one edge
of the
three-sided shaped recess being circumferential and the other two edges of
this three
sided shape corresponding to the inlet and outlet cut-off points when the
edges of the
recesses move past the inlet and outlet ports respectively, the pattern and
orientation of
the two outlets and two inlets alternating around the circumference, wherein
also the
water wheel or turbine like effect of steam entering the valve and
encountering an edge
of the recess assists rotation of the cylinder in a constant direction.


19



8. The rotary valve cylinder of claim 2, and claim 7, is mounted coaxially on
a sturdy
rotating shaft such that;

a, it allows a close fitting but free longitudinal movement of the cylinder
along the shaft,
this being effected by mating shapes of the outer surface of the shaft and
inner surface
of the hole formed in the cylinder, such as is commonly accomplished through
spines,
keys and keyways, and cross sections of polygons both regular and irregular,
with
abutments and locking devices such as screws, pins and the like, such that the
length of
movement of the cylinder along the shaft may be adjustably secured,

b, the shaft extends from at least one end of the cylinder, and generally at
least one at
each end, such extension being secured by rotary bearings, the inner portion
of the
bearing being secured near at least one end of the shaft, and the outer
portion of this
rotary bearing being secured to the valve housing of claim 9,

c, the shaft of claim 8.b is turned at the same speed as the engine of claim
1, the shaft
being connected to the main drive shaft of the engine by a rotary transmission
device
such as gears, timing belts, especially notched belts and pulleys, timing
chains, and the
like, the shaft being turned at the same angular velocity as the main engine
drive shaft,
the rotary transmission device being connected to at least one of the main
engine drive
shafts, the advantage of notched timing belts and pulleys rather than timing
chains and
timing gears being that there is a very smooth action and adjustments in
advancing and
retarding timing may be easily accomplished via jockey pulleys and the like,
and in the
case of timing gears these may be connected to a separate set of spur gears,
bevel
gears, and the like, whereby closer approach of the rotary inlet valve and the
main
engine inlet may be accomplished by one skilled in the art, the second set of
gears, or
second portion of the main gears being mounted on the main drive shaft,
rotating
together but separate from the main engine synchronising gears, whereby uneven
wear
on the main engine synchronising gears is avoided,

d, the moment of inertia of any additional rotating mass connected to one of
the main
engine drive shafts directly in the form of at least one separate portion of
the main
engine synchronising gear wheels, and via the rotary transmission device
including the
rotary valve itself, being balanced by the other main engine rotary piston and
its
synchronising gearwheel having appropriately increased and symmetrically
distributed
mass, whereby rotational acceleration occurs without unbalanced inertial
reactions of
the whole engine,

e, the rotary transmission device of claim 8.c, has rotational adjustment of
at least one
its elements such that equal advancement or retardation of all the inlet cut-
offs may be
effected, examples of such rotary adjustment being those made by minor
rotation of the
rotary mechanism connected to the main engine synchronising shaft, this being
able to
tum slightly and being adjustably secured by grub screws, tapered screws and
bolts,
lock nuts, tapered keys, pins in a set of holes and the like, similar rotary
adjustments
being effected at the rotary transmission component secured to the shaft of
the rotary
valve, and means of changing the length of timing-belt or chain by the action
of





additional rotary components such as jockey pulleys and the like, by adjusting
the
placement of the valve housing with respect to the main engine, whereby
advancing
and retardation of inlet cut-off timing is effected.

9. The balanced rotary inlet cut-off valve of claim 1, includes a valve
housing in the
shape of a hollow cylinder with the ends securely sealed, with at least one
end on the
housing having a circular hole formed at its centre to allow the free but
close fitting
rotation of the shaft of the claim 8 within the housing, the shaft protruding
from the
housing sufficiently to connect to the rotational transmission device and
rotational
adjustments of claim 8.b and 8.c, the valve housing being a hollow cylinder
with internal
diameter allowing free rotation with a close clearance with the grooved
cylinder of claim
3, or recessed cylinder of claim 7, although strict steam tightness not being
necessary,
that function being performed by steam seals associated with protecting the
bearings of
claim 8.b from high pressure steam, and with additional steam seals being
situated at
the outer boundary of the valve housing, at least one end of the usually
closed ended
cylinder housing being able to be removed and re-secured using bolts, screws
and the
like, positioning lugs and keys, gaskets and the processes usually associated
with the
sealing of pressure vessels of this nature commonly known to those skilled in
the art,
whereby the housing can be easily assembled and dissembled for manufacture,
maintenance and repair.

10. Holes for inlet and outlet of steam are formed in the valve housing of
claim 9, with a
predetermined relatively small distance separating the adjacent boundaries of
each of
the inlet and outlet ports, the predetermined distance being such that the
material of
manufacture does not deform under the steam pressure exerted, and such that
the
angles of entry and exit of inlet and outlet port into the valve housing are
suitable for the
material of manufacture, the angle of entry and exit of inlet and outlet ports
lines, firstly
being substantially within the plane that includes a pair of grooves, and
secondly at an
angle to the curved surface of the cylinder that minimizes turbulence, this
later
requirement favouring a shallow angle with a rounded edge, although not
excluding
other angles and other contours, the angle selected substantially matching the
angle of
the groove as the short curves exit the cylinder as described in claim 5.a,
the holes in
the housing being of generally circular shape, and broad enough to extend at
least over
one groove and simultaneously over one partition between grooves, whereby
sliding of
the inlet and outlet cut-off ports relative to the cylinder performs a smooth
transition from
one cut-off setting to another with approximately equal cross section of steam
conduit
being available at all times.

11. The steam powered equal double rotary piston power plants of claim 1 have
a
secondary expansion of steam whereby back-pressure from the inlet of secondary

expansion does not impart back-pressure to the non-driving piston faces of the
primary
expansion, this being effected by placing two early exhaust ports in addition
to the usual
central midline positioned primary exhaust port, one early exhaust port being
in each
side of the primary expansion chamber, with one early exhaust for each rotary
piston,
through which early exhaust ports, steam is routed to secondary expansion,


21




a, the placement of the early exhaust being such that the opening of the port
commences at a point around the periphery of the expansion chamber that is
adjacent
to the leading piston face of the non-driving rotary when the trailing face of
the same
non driving piston has just come into close proximity with the expansion
chamber
housing, thus trapping moderate pressure steam in between the leading and
trailing
faces of the non-driving rotary piston, the pressure of this steam being about
the same
as the steam after primary expansion at the region primary steam input, the
moderate
pressure steam then being vented into secondary expansion after it is no
longer
connected to the primary input region and before that trapped steam is exposed
to the
central primary exhaust,

b, the early exhaust ports being holes in the expansion chamber commencing at
the
point described in 11,a, and extending a suitable small distance and with a
suitable
cross section that is able to vent most of the trapped steam at the moderate
pressure
within the time taken for about one quarter of a revolution of the main
engine,

c, the early exhaust port holes preferentially being of an aerodynamic cross
section and
the holes entering the primary expansion chamber at an aerodynamic contour,
generally
being in the plane of the two rotary pistons, at least initially,

d, the early exhaust port hole entering at a shallow angle to the tangent of
the circular
shape of the primary expansion chamber, whereby the movement of the trapped
steam
is assisted,

e, the interface between the early exhaust port hole and the circular curve of
the primary
expansion chamber having an aerodynamically contoured leading and trailing
edges
suitable to the materials of manufacture and the forces involved,

f, the cross sectional surface area of the conduit towards secondary expansion
being at
least constant, not decreasing, and preferably very slightly increasing to
assist in
transfer of a large volume of moderate pressure steam,

g, and the three dimensional shape of conduit to secondary expansion being an
aerodynamic curve, in at least two dimensions, directed towards the region of
the
central primary exhaust, though not confluent with the primary exhaust, such
that the
early exhaust from both primary expansion rotary pistons is merged via
conduits of
equal length, whereby alternate pulses arrive at the inlet of secondary
expansion,

h, the secondary expansion engine being a low to moderate pressure rotary
engine,
preferably an equal double rotary piston engine of suitable size, although not
excluding
other power plants such as turbines, "Roots" blowers in reverse and
reciprocating
engines,

i, the resulting secondary expansion being either,



22




j, an auxiliary engine, not linked mechanically to the primary expansion,
whereby conflict
between optimal compound expansion of both secondary and primary expansion
with
varying loads is avoided, the secondary expansion engine preferably driving
electrical
generator systems and other ancillaries, or,

k, a compound engine, mechanically linked, in which primary drive systems are
linked to
secondary expansion by sharing the same drive shafts, or via another fixed or
variable
mechanical rotary transmission system,

l, in a mechanically linked compound expansion with each early exhaust being
routed
separately to the secondary expansion engine mounted on the same drive shafts
as the
primary expansion, the routes taken from early primary exhaust to secondary
expansion
inlet being the shortest possible aerodynamic paths, and the phase
relationship
between the rotary pistons of primary and secondary expansion rotary pistons
being
such that the pulse of early exhaust arrives at the secondary expansion inlet
at
approximately the typical time for one of the secondary rotary pistons to
arrive at the
beginning of an expansion cycle, and the raised cam-like portions of the
primary and
secondary rotary piston on the same axle having a suitable out of phase
relationship,
the phase relationship being that which has the driving force of primary
expansion
occurring as much as possible while the secondary rotary piston is non-
driving, and
visa-versa, thus minimising wear on the rotary pistons and associated
synchronising
gears and bearings,

m, the exhaust from secondary expansion system of claims 11, a-11,i, may be
condensed at least partially before being merged with steam from the residual
primary
expansion exhaust, thus reducing reflux from primary expansion exhaust into
secondary
expansion exhaust, although early merging of primary and secondary expansion
exhausts and rapid condensation is not excluded.


12. Steam sealing of the flat surfaces of the expansion chamber of the steam
powered
equal double rotary piston engine of claim 1, substantially being;

a, each rotary piston having two flat faces, each of these flat faces being
fitted with a
single curved flat seat fitted in a curved groove near the periphery of the
flat surface of
the rotary piston, the curve following each rotary piston's two semicircular
arcs of
greater and smaller radii and the two gear tooth profiles that form leading
and trailing
piston faces, the groove containing the seal being formed deep enough to allow
the seal
to be well supported by the sides of the groove, the base of the groove
housing
recesses that fit springs at an appropriate number and positioning around the
seal such
that suitable relatively evenly distributed pressure is exerted on the seal,
the seal being
wider and or deeper at the sharper comers thus withstanding the extra stresses

encountered at these regions,



23




b, each rotary piston having two flat faces, each of these flat faces being
fitted with a set
of straight seal segments or a single polygonal shaped seal, the seal being
either an
irregular or regular polygon, the straight segments alternatively being
shallow curves,
with curvature less than that given by an arc centred on the rotational centre
of the
piston at that point, whereby wear is distributed over a greater region of the
flat surface
of the rotary piston and long term steam sealing is improved, and ease of
manufacture
is assisted,

c, four circular seals being set in grooves formed in each of the two flat
parallel side
surfaces of the expansion chamber engine housing, the seals being centred on
the
rational axis of each rotary piston, whereby leakage of steam down the side of
the flat
surface of the rotary pistons and into the main drive shaft bearings is
reduced,

d, two substantially straight seals in grooves formed in each of the flat
parallel sides of
the expansion chamber, the grooves and seals being orientated radially with
respect to
the axis of each rotary piston, and orientated on a plane including both axes
of both
rotary pistons, the breadth of the seal being at least wide enough to extend
across the
circumferential distance of the suitable gear tooth profile of the rotary
pistons as they
engage at the central point, the seal being broad enough to prevent it
becoming unduly
tilted during the passage of the two piston faces at the central point,
whereby sealing of
the flat side of the housing at the critical central point is improved,

e, the seals of claim 12,a and 12,b, are combined with two circular seals
being joined by
a straight seal, the straight seal being at right angles to the tangent of
circular seals, all
components being in a flat plane, the region of joining being suitable
contoured and the
thickness at this join being suitable to withstand the extra forces imparted
in operation of
the combined seal, whereby more stability is imparted to the central seal by
anchoring
in with the circular seals, without excluding the use of separate straight and
circular
seals with separate spring loaded or keyed supports or the like.


13. The expansion chamber of claim 1 has shallow grooves formed in the flat
and
curved surfaces of the expansion chamber whereby pressurised steam enters
these
grooves and does not perform useful expansion, but rather results in reduced
passage
of steam through the small space between the rotary piston and housing as
leaking
steam encounters greater turbulence and hence encounters greater than usual
resistance, the grooves on the curved portions of the expansion chamber being
substantially parallel with the rotational axis of the main drive shafts and
spaced at
substantially regular intervals around the periphery of the expansion chamber,
and the
flat surfaces of the expansion chamber having similar grooves directed
radially, or at
least substantially at right angles to the axis of rotation of the main drive
shaft and
extending from a diameter a small distance less than the diameter of the
smaller
diameter of the rotary piston until the curved surface of the expansion
chamber, the
grooves on the flat surface generally intersecting with the grooves on the
curved surface
of the expansion chamber, any sealing of the main shaft bearing by additional
seals
having a suitable clearance with the grooves.



24




14. Counterbalancing weight or weights may be placed symmetrically within the
non-
raised half of the rotary piston of claim 1 such that the piston is balanced,
the material of
the weight being more dense than that of the bulk of the rotary piston, such
material
being an alloy of high lead content or the like, additionally or
alternatively, at least one
hole may be formed symmetrically within the raised half of the rotary piston
for the same
purpose.


15. The equal double rotary piston of claim 1 with the elements of primary
expansion,
balanced rotary variable inlet cut-off valve, secondary expansion, and
ancillaries driven
by secondary and primary expansion is arranged and orientated spatially such
that the
sense of rotation, clockwise or anticlockwise, of the primary expansion and
the
associated rotary transmission system driving the main load such as road
wheels in an
automotive application, is constructed to be in the opposite sense of rotation
and on a
substantially parallel axis to the balanced rotary variable inlet cut-off
valve and any
rotary ancillaries driven by the secondary expansion engine, such as
electrical
generators, whereby during acceleration net changes in angular momentum of the
drive
train attached to primary expansion is balanced by net changes in angular
momentum
of a combination of the balanced rotary inlet cut-off valve and the
ancillaries driven by a
secondary expansion engine if one is used, the net changes in angular momentum

within the core mechanism of the primary and secondary expansion being
necessarily
zero due to the balanced geometry of the mechanism as a whole and individual
rotary
pistons as in claim 14, or via other types of balancing as customary to one
skilled in the
art.



25

Description

CA 02761785 2011-11-14
WO 2010/132960 PCT/AU2010/000706

Rotary Piston Steam Engine with Balanced Rotary Variable Inlet-Cut-
Off Valve and Secondary Expansion without Back-Pressure on
Primary Expansion

Contents Page
Title 1
Contents 1
Background 2
Advantages of this Engine, 1-17 3
Explanation of Concept Sketches, 1-11 6
Balanced Variable Inlet Cut-Off Rotary Valve 9
Operating Principle in an Example of Four Inlet
Cut-Off settings 10
Detailed Description of the Balanced Rotary Variable Inlet
Cut-Off Valve, In a Four Cut-Off Setting Example, 1-8 10
Secondary Expansion of Steam in an Equal Double Rotary
Piston Engine - The Problem of Back-Pressure 13
Solution to the Problem 13
Claims, 1 - 15 18
Sketches, 1 11 26 - 36

1


CA 02761785 2011-11-14
WO 2010/132960 PCT/AU2010/000706

Rotary Piston Steam Engine With Balanced Rotary Variable Inlet-Cut-off
Valve and Secondary Expansion Without Back-Pressure on Primary
Expansion

Background
The "equal double rotary piston" mechanism was patented in France in 1832 but
its
potential has never been fully realised because of various problems which are
now
solved with a balanced rotary variable inlet cut-off valve and secondary
expansion
driven by the exhaust of the primary expansion in a manner which does not
impart
back-pressure against the primary expansion.

Sketch 1 teaches the basic geometry of the "equal double rotary piston"
mechanism.
There are two equal disc-like rotary pistons mounted on parallel shafts and
housed
within an expansion chamber closely fitted around the path traced by raised
semicircular portions of the rotary pistons. The raised semicircular surface
of one rotary
piston and the non-raised semicircular surfaces of the other rotary piston
have a dose
approach at the central point of the expansion chamber. The piston "faces"
between the
raised and non raised portion of the rotary pistons are of a suitable gear
tooth profile.
The top of the raised cam-like portion of the rotary piston extends nearly
1800, this long
distance providing good sealing despite absence of piston rings. The two
rotary pistons
are secured on two parallel drive shafts, each shaft being secured to a geared
wheel
external to the expansion chamber. These two equal gear wheels engage and turn
the
rotary pistons in synchrony, at equal speeds but opposite directions.
Pressurised steam,
(or any other working fluid), enters one side of the expansion chamber near
the centre
of the mechanism. This fluid exerts a pressure on the driving face of one of
the pistons,
the pressure being at approximately normal to the plane containing the axis of
rotation
and the radius that passes through the piston face. In other words, the
pressure is
exerted at the near optimal orientation of the piston face, developing near
maximum
possible turning moment from the pressure. The raised portion of the other,
non-driving,
rotary piston forms an abutment. The pressure directed centrally is taken by
the
bearings on its shaft - without any expenditure of energy apart from
frictional losses in
the bearing as it turns. One rotary piston is driving for half a turn, while
the other is
driven, the situation is then reversed for the second half turn - and so on.

The mechanism is slightly similar to a single lobed gear pump operated in
reverse as an
engine. However, because a single lobe would not produce continuous rotation
the
motion is maintained by an external set of gears.

The two pistons are of equal shape, unlike many other attempts at rotary
piston
mechanisms. For this reason the mechanism can be conveniently described,
although
not fully defined, as the "equal double rotary piston" mechanism. Fuller
definition
includes the raised portion of the rotary piston being a circular arc of
nearly 180
degrees, fitting closely within the expansion chamber, as well as the two
rotary pistons
moving in close approximation as they rotate on parallel axles in opposite
directions,
synchronized by gears external to the expansion chamber.

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CA 02761785 2011-11-14
WO 2010/132960 PCT/AU2010/000706

Advantages of This Engine: There are many advantages of the equal double
rotary
piston engine and very few weaknesses. Overall it should be a far better
engine than all
current automotive engines.

1. The rotary pistons continually rotate in opposite directions thus
functioning as
flywheels and so conserving energy very efficiently. The complete absence of
energy
wastage via reciprocating or oscillating movement, even in minor components,
(such as
valves or abutments), is a major energy conserving factor.

2. There is a near 100% power stroke of the "cycle" - in contrast to the 25%
power
stroke of a four stroke internal combustion engine.

3. The resolution of forces in a reciprocating piston and crank means that the
piston and
con-rod act only very briefly in a near optimal orientation to produce
rotation. (The optimal
position is when the piston and con-rod apply all their forces in a straight
line, and when
this is also at right angles to the arm of the crank. This is approximated
only briefly each
cycle, but never fully satisfied in finite sized crank engines.) In contrast,
the equal double
rotary piston engine always applies force to the piston face is at nearly
right angles to the
rotating shaft, (depending on the slope of the gear profile), producing near
optimal turning
moment nearly 100% of the time. It is also vastly superior to the Wankel
rotary mechanism.
4. Converting reciprocating motion into rotary motion via con-rods and
crankshafts also
creates friction as a reciprocating piston has components of forces directed
against
cylinder walls at changing angles. Such friction causing loss of power and
efficiency is
avoided in this particular rotary engine.

5. Also unlike a typical internal combustion engine there is no loss of power
by
induction, pre-ignition compression, and exhaust strokes. Driving of cams and
valves is
also eliminated. Such energy expenditure is often against springs operating in
a non
elastic manner, frequently involves reciprocation, and entails significant
friction.

6. The two rotary pistons turn in opposite senses, clockwise and
anticlockwise, ensuring
that their acceleration imparts no net rotary inertial forces to the housing.
This is an
important advantage in automotive power plants where engine mountings are a
significant part of the power to weight optimisation.

7. Both rotary pistons as well as the inlet cut-off valve are perfectly
balanced and so
produce no vibration at both high and low speed. This reduces the bulk of
engine
mounts, and improves power to weight engineering generally.

8. The mechanism is a positive displacement engine, not a turbine. This
results in good
acceleration from a stationary position against a load - as is required
especially in
typical automotive applications. Turbines are very poor in accelerating from a
stationary
position against a load. The "equal double rotary piston" mechanism is not an
orbital
engine, vane engine, or a Wankel engine - which despite being positive
displacement

3


CA 02761785 2011-11-14
WO 2010/132960 PCT/AU2010/000706

rotary engines, all have one or more major problems especially in automotive
applications.

9. With a constant direction of rotation and near constant angular velocity
during a given
cycle, there is very little friction and wear. Modem bearings, seals and
timing gears
have all been highly engineered over generations for great durability and
performance
and are easily utilised in this new setting.

10. The long curved surface of the raised portion of the rotary piston and the
long
distance in which it is in close proximity to the housing ensures that very
little steam can
leak between these surfaces, despite the absence of piston "rings". The two
surfaces
have their maximum length in close approximation when it is most needed, that
is when
the pressure is at a maximum - at the beginning of expansion. The flat faces
of the
rotary piston have seals which prevent steam escaping past the side of the
raised
portion of rotary piston and also from escaping through the main drive shaft
bearings.
11. With modem accurate manufacturing techniques there will be a very small
constant
clearance between the two rotary pistons at the central point where there is
tangential
approach of the two rotary pistons. This allows a small amount of steam to
escape
between the rotary pistons at the central point. This steam is kept within the
sealed
system and exits with the exhaust steam, which is then condensed and re-used.
This is
the only weakness of the whole design. It is amply compensated for by the many
advantages of this rotary engine over reciprocating and other rotary engines.

12. Both rotary pistons function as both piston face and abutment in one solid
robust
member. This important fact distinguishes the mechanism from the vast majority
of
other rotary piston mechanisms. Many other rotary piston designs have
separate, often
small, moving and therefore relative flimsy abutments. Spreading out the wear
evenly
over long and uniformly curved surfaces in dose proximity to its adjacent
surface
distinguishes the mechanism from another common weakness of many other rotary
piston designs. The balanced rotary inlet cut-off valve is also a very robust
simple
design with excellent durability.

13. Since it is an external combustion engine, burnt fuel residues do not
enter the
expansion chamber and produce contaminate or deposits - unlike internal
combustion
engines. Oil for bearings and synchronising gears is thus kept clean,
resulting in low
maintenance and improved longevity of the engine.

14. Properly controlled external combustion can produce less atmospheric
pollution and
allows a wide choice of many different fuels. Fuels used to produce steam may
include
traditional petroleum based fuels such as gasoline, kerosene or L.P.G.
(Natural Gas).
However these fossil fuels are contributing to net increases in atmospheric
carbon
dioxide, global warming and adverse climate change. More environmentally
responsible
fuels are being developed. These include renewable sources such as (second
generation) ethanol, and algal oil. Less ideal fuels are first generation
ethanol and
vegetable oils such as canola. Hydrogen can be used as an external combustion
fuel

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generated from a variety of intermittent energy sources such as wave, wind and
solar
power or constant sources such as geothermal energy. However the more direct
combustion of second generation biofuels is a more direct, better, option than
hydrogen.
15. Whatever the ultimate energy source the engine uses high pressure steam.
For at
least the last 30 years technology for rapid steam production in sufficient
pressures and
quantities for typical automotive use can be generated in about 45 seconds.
Modem
steam generators for automotive use are compact, light weight, safe and
reliable.
Regulation of steam for automotive use is also now very well established.
Neither of
these two areas of technology will be discussed in detail in this patent
application.

16. The equal double rotary piston steam engine is simple, compact, with few
moving
parts and relatively inexpensive to manufacture. This is demonstrated by the
fact that
the prototypes of the rotary piston engine were produced in a backyard garage
which
had only a small lathe, a drill press, hand tools and air compressor. Even
allowing for
the cost of a steam generator, production costs would be cheap compared to
those of
internal combustion engines.

17.The suitability of a steam powered equal double rotary piston" power plant
to
automotive applications is such that it would be possible to dispense with a
clutch and
gearbox, as some less efficient reciprocating piston steam vehicles have
successfully
done in the past. The weight saving of eliminating clutch and gear box would
add to the
power to weight efficiency of the power plant and reduce manufacturing and
operating
costs even further. It might be possible to eliminate other portions of the
transmission
train by having separate smaller equal double rotary piston engines directly
driving each
driven wheel. However this advantage would be offset by the need for
relatively greater
total thermal insulation of at least two engines, and by extra measures taken
to protect
the engines and pressure conduits from more proximity to road vibration.

Consequently we are of the opinion that this engine as described in the
present patent
application has the potential to make the four-stroke internal combustion
engine
obsolete in many situations. Automotive applications include heavy-duty long-
distance
road transport, light and commuter transport, as well as rail and marine
transport, and
possibly even air transport! ( Regarding air-transport, the present engine and
modem
steam generators would be far more efficient than the very successful
reciprocating
piston steam engine propeller biplane of William and George Besler, flown in
April 1933.
The original newsreel is at www.youtube.com/watch?v=nw6NFmcnW-8.) Stationary
applications include large scale electricity generation and small scale
combined heat
and electric power generation. Farmers could produce their own electric power
using
their own fuel as virtually any combustible fuel can be used in a furnace to
produce
steam. Portable units could also produce electricity or operate pumps for
pneumatic or
hydraulic equipment. Many tools, including compressed air machines often used
in the
mining industry, could be easily adapted to the "equal double rotary piston
engine".
Many other industrial processes could use steam powered equal double rotary
piston
power plants, making a more direct and hence efficient use of local energy
sources.



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Explanation of Concept Sketches

Sketch 1. This sketch shows elevation and sectional elevation through the
rotary
pistons. It shows the engine at the transition from rotary piston N 2 driving
to the other
piston (N 1) driving. An expansion chamber is formed by the housing and the
pistons.
The leading surface of the elevated portion of rotary piston has a suitable
gear tooth
profile shape, this curved face forming the piston face. The engine will
always turn
rotary piston 2 clockwise and rotary piston 1 anti-clockwise. The purpose of a
gear tooth
profile, (such as involute or other suitable curves), on the piston face is to
minimise
steam escaping during the brief transition from one part of the cycle to the
next,
because the small gap remains constant till they separate. The non-driving
rotary piston
maintains an abutment at the rear of the expansion chamber, against which
steam
pressure applies force to the leading end of the piston face which is in its
driving cycle.
Rotary piston 2 will continue to drive till its trailing gear profile face
completes its
transition and the other rotary piston becomes the driver. This will occur
alternatively for
each rotary piston after it turns 1800. Thus power is delivered alternatively
for half the
cycle by one rotary piston, then the other, so that the driver becomes the
driven gear
and the driven gear becomes the driving gear at each transition. What is said
in respect
of one rotary piston in any of the following sketches, applies equally to the
other rotary
piston when it is in the equivalent position, bearing in mind that they turn
in opposite
directions. Despite there being two rotary pistons the engine should not be
regarded as
a twin cylinder engine, because one rotary piston will not work without the
other.
The two rotary pistons are synchronised by gears on each rotary piston shaft.
These
gears have the same pitch circle diameters as the rotary pistons. That is they
share the
same mid-point diameter of the smaller and larger diameters of each rotary
piston.
Following are brief descriptions of the various phases of the cycle, much of
which is
self-explanatory in the diagrams.

Sketch 2. Rotary piston 2 has just finished its power stroke and rotary piston
I is about
to start its power stroke.

Sketch 3. Note that; a, inward pressure on diameters produces no turning
motion and,
b, there are three gear profile faces open to driving pressure - the force on
the two
faces of rotary piston 1 are balanced giving no net driving force, while the
unbalanced
forces on rotary piston 2 produce clockwise rotation of that piston.

Sketch 4. Rotary piston 1 has passed the middle of its power stroke and has
almost
ceased exhausting its previous power stroke. Rotary piston 2 has started to
exhaust.
Sketch 5. This shows a sectional view of a balanced variable inlet cut-off
rotary valve in
an example with four predetermined inlet cut-off settings, see pp.9-12. Both
the number
of cut-offs, (not just four), and the cut-off ratios, can be chosen to suit
specific
applications.

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Sketch 6. This isometric sketch illustrates the double sided nature of the
balanced
rotary valve. The full three dimensional nature of the solid cylinder with
grooves formed
is not illustrated, merely the outer edges of the grooves on the surface of
the cylinder.
Again a four cut-off setting is shown as an example.

Sketch 7. This sketch shows an example of the rotary inlet cut-off valve in
relation to
the engine. It illustrates an example with equal lengths of steam travel at
all equivalent
stages, from bisection of inlet to inlet cut-off valve, through inlet cut-off
valve itself, and
exit from inlet cut-off valve before merging and then entry into the expansion
chamber. If
a toothed timing belt it used to directly connect the main engine drive shaft
and inlet cut-
off, a similar geometry may be used. In simple pulley systems rotation takes
place in the
parallel planes - however other rotary transmission systems allow non parallel
paths.
For example, one may use a system of bevel gears between the main engine drive
shaft and the inlet cut-off valve shaft, allowing a closer approach of the
inlet cut-off valve
to the entry to the expansion chamber. The axis of rotation of the rotary
valve may be at
right angles to the axes of the main drive shafts, passing through the central
point, and
in the plane of rotation of the rotary pistons. In this system the tow paths
of steam have
the same contour in the central region - an "S" shape, and a mirror image "S",
with cut-
off occurring at the centre of the "S" shape. This is not shown in the
sketches.

Sketch 8. This sketch shows two possible arrangements of early exhaust port
for
extraction of trapped steam for secondary use, see pp. 13 -12. Secondary use
may be
either in a mechanically linked "compound engine", or in a non-mechanically
linked
"auxiliary engine". An auxiliary engine may be used to generate electricity or
drive other
ancillaries of automotive use etc. Note the aerodynamic path taken by the
steam en-
route to secondary expansion at a region equally favourable to steam routed
from both
rotary pistons, namely a midline region near the initial primary expansion
exhaust.
Sketch 9. The sketches 9, 10, 11 show several ways to seal rotary pistons at
their flat
faces. These approaches are additional to those in our patent W02006102696
(Al),
published 2006-11-16, priority AU20050201741 20050427.

The curved planar seal is fitted in a groove around the periphery of the flat
surface of
the rotary piston. The groove is deep enough to allow the seal to be well
supported by
the sides of the groove. At the base of the groove are recesses that fit
springs of an
appropriate number and positioning around the seal, such that suitable
relatively evenly
distributed pressure is exerted on the seal. The seals could be wider at the
sharper
comers to bear the extra stresses encountered at these regions - this last
feature is not
shown in sketch 9. A series of straight seals or a single polygonal seal with
straight
segments may be used instead of a curved seal. The seal may be either an
irregular or
regular polygon and these straight segments may be replaced by shallow curves,
with
curvatures less than that given by an arc centred on the rotational centre of
the piston at
that point. The advantage of straight or slightly curved segments is that wear
is
distributed over a greater region of the flat surface of the rotary piston.
Straight

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segments may be less expensive to manufacture. The dashed line of sketch 9
shows an
example of one possible arrangement of straight segments.

Counterbalancing weight or weights may be placed symmetrically within the non-
raised
half of the rotary piston so that the piston is statically balanced. The
weight would be of
a material denser than that of the bulk of the rotary piston, possibly
tungsten or a lead
alloy. In addition or alternatively, at least one hole may be formed
symmetrically in the
raised half of the rotary piston for the same purpose - (not illustrated on
sketch 9).

One may discuss at this point the balancing of the power plant as a whole. The
primary
expansion, balanced rotary variable inlet cut-off valve, secondary expansion,
and
ancillaries driven by secondary and primary expansion all rotate and ideally
this rotation
should be balanced, especially in acceleration. The core mechanism of primary
expansion is inherently balanced, but the associated rotary transmission
system driving
the main load such as road wheels is not balanced. Likewise the "balanced"
rotary
variable inlet cut-off valve is not balanced with respect to dynamic angular
momentum -
merely in the balancing of forces on its bearing, and static balancing.
Similarly, any
rotary ancillaries driven by the secondary expansion engine, such an
electrical
generator, are usually unbalanced during acceleration. The spatial arrangement
of all
these systems which accelerate together can be arranged such that net changes
in
angular momentum mostly cancel out. The component with greatest unbalanced
angular momentum would be the drive train attached to the primary expansion -
that is
the driving wheels etc. This major source of in-balance during acceleration
may be
offset by arranging the sense and direction of rotation of the rotary inlet
cut-off valve and
any ancillaries driven by a secondary expansion engine. A dual electric
generator with
clockwise and anti-clockwise rotors would be balanced, just as is the double
equal
rotary piston engine itself.

Sketch 10. This sketch illustrates more approaches to seating. Firstly,
circular seals
may be set in grooves in the engine housing to prevent leakage of steam down
the side
of the flat surface of the rotary pistons, and into the main drive-shaft
bearings. Secondly,
improved sealing of the flat side of the housing at the central point may be
effected by a
second seal, broad enough such that the circumferential distance of suitable
gear tooth
profile is about the same as, or just less than, the breadth of the seal. This
prevents the
seal becoming unduly tilted during the passage of the two piston faces at the
central
point. Further improved seating is via shallow grooves in the flat and curved
surfaces of
the expansion chamber. Steam enters these grooves and does no useful
expansion.
However turbulence on entering these grooves means further passage of steam
through the small space between rotary piston and housing encounters more
turbulence
and resistance to leakage via that path. Whether this effect is beneficial and
does not
merely increase resistance with little reduced leakage, must to be decided
empirically.
Sketch 11. This sketch is very similar to sketch 10, except that straight, or
at least less
curved, segments are used for the seals as in sketch 9.

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Balanced Variable Inlet Cut-Off Rotary Valve.

See sketches 5, 6, 7 for a visual presentation of the basic concept. There is
an absence
of any form of inlet cut-off in nearly all prior art of steam powered equal
double rotary
piston engines. This was probably an important factor contributing to this
exceptional
mechanism not becoming widely used engine long ago. Prior art disclosing any
form of
rotary variable inlet cut-off for an equal double rotary piston engine is
unknown to us.
Furthermore, the proposed valve is balanced, both statically balanced, and
balanced
with regard to the pressure exerted by steam on the bearings of the rotary
valve.
Typical automotive power plants have rapidly varying loads and widely varying
speeds.
It is indispensable to quickly and smoothly change between two, three, four,
or more
inlet cut-offs to use the most appropriate amount of steam to balance power
and
economy. This design is capable of rapid and smooth changing between
potentially a
large number of inlet cut-off settings. Consequently we believe it is
especially suitable
for automotive application. In some stationary situations, such as electrical
power
generation with its slowly varying loads, only one or two inlet cut-off
settings may be
required. The valve is simple, easy to manufacture, effective, robust and
durable. For
these reasons we believe that this design and application to be novel and very
useful.
The importance of inlet cut-off to allow fuller steam expansion was realised
by steam
engineers in the 1830's. Without any inlet cut-off a full head of steam can
push a piston
slowly against a large load, and at the end of the stroke exhaust steam may
still be near
full pressure. Exhausting high pressure steam is wasting energy.

A typical modem automotive steam generator can produce steam at least 20 times
atmospheric pressure. Even in a fast engine operating against a small load it
would be
very inefficient to allow the steam to expand only 10 times in producing power
before it
exhausts to the atmosphere. One possibility would be doubling the length of
primary
expansion, but this is an inefficient way to extract the energy from steam
already
expanded 10 times - as most efficient energy transfer, or work is done early
in
expansion. A better way is cutting off the inlet steam part-way into the
expansion, thus
allowing a more full expansion than with full pressure steam being applied to
the piston
throughout the expansion. The rotary inlet cut-off valve allows the steam to
enter the
rotary engine at the same position at the start of the "power stroke" of each
rotary piston
(ie. twice in 360 revolution) but cuts off the steam at approximately 10%,
30% or 60%
of each power stroke, or it may allow the steam into the engine continuously
for 100% of
the cycle. Bear in mind that there are two rotary pistons in the engine. One
drives for
half one revolution (i.e.180 ), and then the other rotary piston drives for
half one
revolution. So in one 360 revolution of the engine each rotary piston in turn
drives 180
while the other exhausts - giving a near continuous power stroke.

This "balanced variable inlet cut-off rotary valve" was initially designed for
use with the
equal double rotary piston rotary steam engine to improve efficiency. However
the valve
may also be used in other applications.

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Operating Principle in an Example of Four Inlet Cut-Off Settings:

a. Consider for example a 10% "economy" setting. This allows steam into the
engine for
approximately 10% of the power stroke of each rotary piston, allows
approximately 90%
of the power stroke for the steam to expand and achieve near maximum energy
efficiency. The time taken for opening and closing of the valve would probably
reduce
the optimally efficient valve operating time to about 85%.

b. A 30% setting allows steam into the engine for 20% more of the power stroke
than a
10% setting, but it has also 20% less of the power stroke in which to expand
before it
completes the power stroke. This means that more steam has entered during the
power
stroke, but it has had less of the power stroke in which to expand and achieve
its work
potential. This gives more power at the expense of economy.

c. A 60% setting, for the same reason allows steam to enter for 50% more of
the power
stroke than the 10% setting - but is very wasteful of fuel. This would be best
used only
for short periods under extremely large load conditions such as climbing a
steep hill.

d. In automotive settings, if a forward-neutral-reverse mechanical gearbox and
clutch is
used, then for a cold start the 100% setting would be selected allowing steam
to enter
the engine continuously for warming-up the engine quickly - while allowing the
engine to
rotate in neutral gear. Also when the engine is turned off, even if the engine
is still hot,
to restart the engine the drum valve would need to be set in the 100% position
for the
engine to start, because in other settings the valve may stop in a closed
position.
Detailed Description of the Balanced Rotary Variable Inlet Cut-Off Valve
(In a Four Cut-Off Setting Example)

1) The rotary valve has a rotating cylinder on a shaft inside a sealed
cylindrical bore
housing. The inner cylinder turns at the same rate as the rotary pistons in
the engine.
This cylinder has a minimum of clearance with the bore with no metal-to-metal
contact.
The cylinder is keyed or splined to the shaft and can slide along it. It has a
groove cut
right around the circumference of the rotating cylinder (in the 100% setting),
so that
when this groove is aligned with the steam entry and exit ports in opposite
sides of the
cylinder bore, it does not inhibit the continual flow of steam through the
valve.

The cylinder has three (or more) other grooves of different lengths around the
circumference of the rotating drum running parallel to the continuous (100%)
groove
and equally spaced along the drum. The drum may be moved along the bore so
that the
groove of choice may be aligned with the steam entry and exit ports. The start
of each
of these grooves are in-line, and are timed by a toothed-belt drive or gears
such that to
open when the engine rotary pistons pass the engine inlet port.

As indicated, the grooves are of different lengths; for example 10%, 30% or
60% of half
the circumference of the drum. With respect to the grooves, the rotating drum
is double-


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sided. Equivalent grooves are formed in line with these grooves on the other
side of the
drum so that in one revolution of the valve drum, two grooves of the same
length will
pass a given point. Consequently, in one complete rotation of the grooved
cylinder, as
the cylinder rotates and the start of the groove passes the entry port of the
valve, it
allows steam to pass through the groove and out the exit port of the valve
into the
engine for the duration of the groove. When the rear end of the groove passes
the entry
port it cuts off the steam flow for the remainder of the half turn.

The same process happens simultaneously on the other side of the valve because
there
are two sets of grooves on the drum and an entry-exit port on both sides of
the valve
cylinder. This is repeated twice in one rotation of the valve and engine. The
steam
supply line is divided to serve both valve inlets and the two exhausts unite
before the
steam enters the engine inlet port. Thus in one rotation of the valve
cylinder, the steam
enters and exits the valve twice for a short period depending on the cut-off
ratio chosen.
2. Since the steam conduit is divided and enters the valve at opposite sides,
(and joins
again before entering the engine), the force of steam pressing on one side of
the
rotating drum is balanced by an equal force on the other side. This should
result in long
life of the rotary valve bearings. The valve drum is neat fitting but does not
touch the
bore of the cylinder in which it turns. Thus there is no friction except in
the bearings and
seal, and little energy needed to drive A. This solves a friction and wear
problem often
associated with rotary valves, (especially rotary valves in internal
combustion engines).
3. Since the incoming steam will move in the same direction that the rotating
drum
turns, initial impact of steam pressure on the start of the groove acts like a
turbine,
assisting rotation of the drum. This could result in little, if any, effort
being required by
the timing device such as a toothed belt drive to turn the valve - assisting
energy
efficiency.

4. Since the rotating drum does not touch the cylinder bore, there will be
some leakage
around the drum and this will result in the interior of the valve being
pressurised, and in
a small amount of steam continuing through into the engine when the valve
doses. This
will not be a problem as the overall system is sealed, and the leakage of some
steam
into the engine will only contribute positively to driving the engine -
smoothing out the
pulse of inlet cut-off steam.

The purpose of the inlet cut-off valve is to produce a "pulse" of steam for
the duration of
the valve setting, even though it will not fully stop the flow of steam when
the chosen
groove closes. Unlike reciprocating engines in which a leaky valve results in
energy
complete lost, leakage past this inlet valve is merely a small amount of steam
entering
the cylinder without inlet cut-off, and is not wasted, although less efficient
than steam
used with inlet cut-off.

5. The valve receives steam and operates only when the engine is in drive or
warm-up
mode. Movement of the drum along the cylinder bore when choosing a different
mode
will not be inhibited because end-pressure caused by steam trapped at either
end of the

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hollow cavity of the valve housing will be equalised by vents through the
rotating drum.
Instead of the swinging arm selector as indicated in sketch 5, alternatively a
rack and
pinion may be used to move the yoke and slide the drum along its shaft.
Different types
of bearings and seals may be used.

6. When changing cut-off settings, because the steam entry and exit ports of
the valve
are wider than the division between the drum grooves, the next groove starts
to open
before the current one closes. Consequently there is no dead-spot between cut-
off
settings. Combination of two adjacent settings could in effect produce an
intermediate
setting between them - effecting a smoother change of effective cut-off.
Changing of
cut-off settings should be smooth enough to not require the use of a clutch.

7. Rather than a fixed number of discrete inlet cut-off settings, a
continuously variable
inlet cut-off may be accomplished by removing the partitions between the
adjacent
grooves, resulting in a pair of three-sided broad recesses on the surface of
the cylinder.
The comers of the pair of three sided shapes would touch at two of each
triangle's
comers if a continuous groove is included, i.e. in a 100% cut-off setting.

In this continuously variable valve pressurised steam would not be as
effectively
confined to the groove of the path as with discrete groove channels, but the
fluid flow
across the recess would still be predominantly in a 2 dimensional curve
joining the inlet
and outlet ports. This continuous cut-off valve would have increased
turbulence
compared to a valve with a discrete number of grooves. However the advantages
of
continuously variable inlet cut-off may outweigh this disadvantage in
practice.

The shape of the three sided recess may be a (straight edged) triangle wrapped
around
a cylinder for simplicity, but a curved edge, (or edges), could be designed to
advantage.
For example one, may compensate for the non-linear movement of the yoke with
respect to the constantly varying arc through which the simply hinged
actuating lever or
handle is turned, as shown in sketch 5. Alternatively, one could design
suitable changes
in variable cut-off that correspond to empirically determined typically useful
changes in
inlet cut-off during acceleration for the application for which the engine is
designed.

8. This rotary inlet cut-off valve does not alter the mechanical advantage of
the engine
and transmission. However after optimisation of an automotive system, it may
be
decided empirically whether inlet cut-off will serve most of the purposes of a
mechanical
gearbox, or if it is better used in conjunction with a typical gearbox. In
this later case, the
variable inlet cut-off would serve mainly for energy efficiency not mechanical
advantage.
Changing inlet cut-off alters the power and economy but not the ratio of
engine to wheel
revolutions. Since the engine is capable of very high revolutions it will
usually need to
be geared down, even if a typical variable ratio gearbox is not used. The
ratio of gearing
depends on the size of the road wheels, maximum speed of the car and power
needed.
This in turn determines the magnitude of the engine capacity, steam generator
size, fuel
supply, etc.

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Secondary Expansion of Steam in an Equal Double Rotary Piston Engine
- The Problem of Back-Pressure.

Another important improvement in the equal double rotary piston engine relates
to
designing a second engine that uses the low pressure exhaust steam from
primary
expansion without imparting back-pressure onto the non-working faces of the
pistons
involved in the primary expansion. If one merely places the input to a
secondary engine
at the centrally located exhaust port of the primary expansion there will be a
pressure
build-up in the exhaust region of the primary expansion that exerts back-
pressure on the
non-working gear profile face of the rotary piston. Any energy gained by the
secondary
expansion would be at the expense of energy lost from the primary expansion.
Note that
with reciprocating steam engines one can simply use exhaust steam for
secondary
expansion because there is an exhaust valve that closes after primary
expansion such
that back-pressure cannot be exerted back into the primary expansion after
this exhaust
valve has closed. This is an outstanding problem to be solved with the equal
double
rotary piston mechanism, namely, to determine a simple means of including a
secondary expansion of primary exhaust steam without imparting back-pressure
to the
primary expansion. Introducing a new separate exhaust valve on the primary
expansion,
as occurs in reciprocating engines would be one solution - but an inelegant
solution
involving several additional components, friction, possibly reciprocation
losses and cost.
Solution to the Problem.

By careful reference to sketch 8, one can observe that the raised portion of
the rotary
piston 2 has just closed off steam entering its half of the expansion chamber,
and that
rotary piston I has just begun its power stroke. The volume of partly expanded
steam
between the leading and trailing face of rotary piston 2 is effectively seated
and remains
constant for the rest of the rotation, almost a quarter of a turn - until the
leading face of
rotary piston 2 passes the exhaust port. For this period the steam trapped in
this
cylinder cannot expand and does no work. It neither contributes to turning,
nor does it
hinder it. This is a feature that can be used to advantage. While this fixed
cavity
remains, irrespective of its position, its trapped steam can exit through an
early exhaust
port into secondary expansion. Once the leading face of rotary piston 2 passes
the
usual central exhaust port, remaining steam will exhaust out through it and is
not used.
The residual pressure in the central exhaust outlet of the primary expansion
would be
higher than the exhaust pressure of the secondary expansion. Ideally it would
require
two condenser systems. The condenser for the secondary exhaust would be
designed
to operate at a lower pressure than the condenser from residual primary
exhaust.
Merging a high pressure condenser system with a low pressure system would
unhelpfully impart some back-pressure onto the lower pressure secondary
expansion.
However, since there would not be a great difference between the two exhaust
systems
one may merge the two exhaust systems after some initial separate condensation
brings both pressures quite low, hence quite close, after which one may have a
final
combined condenser. Alternatively, once could rely on having such efficient
and rapid

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condensation of an early merged single condenser that the negative pressure of
efficient condensing simply draws in steam from both primary and secondary
exhaust
without putting significant back-pressure on either exhaust. The end result
would be that
combined extra power from both primary rotary pistons would supply steam to a
secondary expansion for almost half of the primary expansion's cycle. This is
a very
considerable advantage in reclaiming of thermal energy into mechanical energy,
energy
that would otherwise be lost out an exhaust or into a condenser.

Expansion Engines: "Compound Engine" (Mechanically Linked) and
"Auxiliary Engine" (Non-Mechanically Linked)

The steam available for secondary expansion could be routed to secondary
expansion
chamber similar to the primary expansion chamber that is mechanically linked
to the
primary expansion giving "compound expansion", or possibly to a separate
"auxiliary
engine" that is not mechanically linked to the primary expansion. A fixed
mechanical
link between primary and secondary expansion such that both expansions drive
the final
drive shaft involves choosing the best compromise ratio of primary and
secondary
expansion. However this optimal ratio varies with varying load since how much
steam
expands at a given speed of revolution depends on how much force it is working
against. Any fixed ratio is necessarily a suboptimal compromise when there are
greatly
varying loads and speeds as is typically encountered in automotive
applications.
Varying the linking ratio via a highly variable gear box coupling primary and
secondary
expansion would be a feasible, but impractical approach. Therefore we believe,
especially in an automotive setting, that a separate, auxiliary engine is
possibly the best
option. The separate auxiliary engine can be used to generate electricity to
charge
batteries for numerous ancillary uses in a fully developed automotive vehicle.
Instead of
the secondary expansion being performed by an equal double rotary piston
engine, with
or without inlet cut-off, one could use a turbine, a "Roots" blower, "gear
pumps engine,
or even reciprocating piston engine. However the many advantages of the equal
double
rotary piston engine make it the best option.

With an auxiliary engine, the placement of the inlet for secondary expansion
must be in
the plane midway between the two main drive shafts of the primary expansion.
To take
advantage of the inertia of the steam trapped between the raised portions of
the rotary
pistons one would route the exhaust destined for secondary expansion through
an
outlet substantially at a tangent to the primary expansion chamber at the
predetermined
point. A gradually expanding cross section of conduit assists forward movement
of
steam. The shallow angle of exhaust take-off, and aerodynamic contours towards
the
central plane described above necessarily favour a secondary expansion input
near the
exhaust port of the residual primary expansion. There could be a small
deviation away
from the plane of containing rotary pistons to allow both a residual primary
exhaust
separate from, yet adjacent to, the secondary expansion inlet. However it is
probably
more important to keep the inlet for secondary expansion less deviated, and
preferentially deviate the path of the residual primary exhaust. The higher
pressure,
higher temperature residual primary exhaust could be used to steam jacket the

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secondary expansion or perform other energy regeneration processes for the
secondary
expansion.

With a dual (primary and secondary) expansion chamber system with two pairs of
rotary
piston operating out of phase, the pistons continually turn both main drive
shafts. In'this
situation the optimal ptacemenf of the secondary expansion inlet would be
between the
two residual exhaust outlets, these outlets passing one on each side of the
secondary
inlet before merging.

Since the secondary expansion steam comes in two pulses per rotation of the
primary
expansion rotary pistons, the secondary expansion need not be merged into a
single
steam flow, but rather each pulse could be synchronised and routed to a
secondary
expansion inlet that is substantially tangential to the direction of steam
flow that is
optimal for inlets at each side of the secondary expansion rotary pistons
respectively.
The shorter the distance between the secondary expansion take-off and the
secondary
expansion inlets, the better. This implies that secondary expansion ought to
have its
inlet posts near the exhaust ports of'the primary expansion. This also implies
that, if the
axefs of the two expansion systems are parallel, as would be a compact and
hence
thermodynamically advantageous arrangement, then the secondary engine would be
"upside down" (with respect to the primary expansion), and its direction of
rotation the
secondary expansion would be opposite to the primary expansion, (i.e.
clockwise
compared to anti-clockwise). This has advantages in minimising vibration, and
reducing
reactive forces due to changes in rotational inertia. Other secondary inlet
placements
with other orientations of multiple chambers can be generalized from the above
examples by one skilled in the art.

Alternatively, with secondary expansion via a compound, (that is mechanically
linked
expansion), then all the primary exhausts destined for secondary expansion
would also
ideally take paths of equal lengths and shape before converging symmetrically
at the
inlet for the secondary expansion. The alternating nature of exhausts from
each one of
the pair of rotary pistons would allow a steady series pressure pulse inputs
into the
secondary expansion giving smooth operation, which as with the non-liked
secondary
expansion described previously, may be routed and synchronised to give optimal
separate inputs to each of the secondary expansion inputs respectively.

With a compound engine each early exhaust may be routed separately and
individually
to the secondary expansion engine which would be generally mounted on the same
drive shafts as the primary expansion. In this situation, the routes from
early primary
exhaust to secondary expansion inlet take the shortest possible aerodynamic
paths,
and it would be favourable to have the two engines mounted near each other and
parallel. The phase relationship between the rotary pistons of primary and
secondary
expansion rotary pistons would ideally be such that the pulse of early exhaust
arrives at
the secondary expansion inlet at approximately the typical time for one of the
secondary
rotary pistons to arrive at the beginning of an expansion cycle. The optimal
time may
vary slightly as with all compound expansion depending on the load. In
practice there



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would be only a slight difference in phase between the two engines as mostly
steam
travels very fast, except under extremely large loads.

If the primary and secondary expansions are mounted on the same pair of axles,
then
the radius of the rotary pistons would have to be the same, and so the
increased
volume of lower pressure steam for secondary expansion would have to be
catered for
by thicker disc-like rotary pistons, with pistons faces less square and more
rectangular
in relative cross section. There would be limits to the rectangular ratio of
such narrow
expansion chamber spaces in terms of efficient expansion due to fluid
dynamics.
Therefore one may consider secondary expansion in a compound engine that is
mechanically linked by a gear train, not simply linked by being on the same
axle. The
secondary expansion's pair of axles could out-flank the primary expansion
axles and
engage with the primary axles via simple parallel gears, bearing in mind that
odd
numbers of gears in a gear train reverse the sense of rotation, (i.e clockwise
to
anticlockwise), and visa versa for even numbers of gear wheels in a gear
train.
Consequently entry into the secondary expansion may be at the "top" or
"bottom" of the
primary expansion depending on the number of gears in the gear train.

With secondary expansion in compound engines that are mechanically linked by
having
primary and secondary expansions on the same axels, there is a simple means of
reducing wear on the external, synchronising gears and main axel bearings.
Consider
on sketch 8, the pulse of steam destined for secondary expansion from the
primary
rotary piston 2. If this pulse of steam is routed to secondary expansion
mounted on the
same axel as rotary piston 2, then it would advantageous from a wear
minimisation
perspective, to have this pulse of steam timed such that the secondary
expansion of
rotary piston 2 is driving whilst primary expansion of rotary piston in non-
driving. This
can be accomplished by routing the secondary expansion steam from rotary
piston 2 up
towards the inlet region of primary expansion, and at the same time having the
raised
cam-like portion of secondary expansion rotary piston 2 being 180 out of
phase with
the raised cam-like portion of primary expansion rotary piston 2. Having the
raised cam-
like portions of primary and secondary expansion on opposite sides of the same
axel
would have some advantages in balancing, but complete balancing would still
need
balancing of both primary and secondary rotary pistons individually.

Similar principles can be applied if one chooses to route the secondary
expansion
steam from rotary piston 2 to an inlet for secondary expansion adjacent to the
exhaust
region of primary expansion. This may be for the purpose of having as short as
possible
path for steam destined secondary expansion before secondary expansion began.
In
order to keep the same clockwise rotation of rotary piston 2, the raised cam-
like portion
of the secondary expansion rotary piston 2 would be about 900 out of phase -
as can be
understood by careful consideration of sketch 8.

Similarly arrangements in which the secondary expansion steam from rotary
piston 2
(on the right in sketch 5), crosses over to the side rotary piston 1 (on the
left in sketch
5), may be constructed by one skilled in the art. This may be for the purpose
of
minimising wear by evening out the driving forces on both rotary piston axles,
and

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simultaneously minimising the length of steam conduit in routing to secondary
expansion, and also by having as aerodynamically smooth conduits as possible.
There are many possible geometries allowing various secondary expansion rates
and
speeds, varying radii and cross sectional areas of secondary expansion piston
faces,
and varying positions of the secondary expansion axels, (generally parallel
to, but not
necessarily coplanar with the primary expansion axels). These variables may be
optimised for the final application. In general, mechanically linked, i.e.
compound
secondary expansions, are best suited for relatively constant loads, or at
least slowly
varying loads, so that fine tuning of these variables for optimal energy
performance can
be achieved. Stationary engines, especially large electric power generation
plants, (and
possible large marine applications), rather than automotive land transport
power plants,
are probably the best setting for compound engines given their relatively
slowly
changing loads, and also given the irrelevance of the extra weight of the
additional
mechanisms in stationary engines. Justifying the extra complications for
relatively small
energy efficiencies is possible in large electric power generating plants
because
relatively small improvements of energy efficiency become significant savings
due to the
very large amounts of total energy conversion involved.

17

Representative Drawing