Note: Descriptions are shown in the official language in which they were submitted.
CA 02701481 2010-04-27
HIGH EFFICIENCY DUAL CYCLE INTERNAL COMBUSTION ENGINE WITH
STEAM POWER RECOVERED FROM WASTE HEAT
This application claims priority based on United States Patent Application
12/387,113 entitled "HIGH EFFICIENCY DUAL CYCLE INTERNAL COMBUSTION
ENGINE WITH STEAM POWER RECOVERED FROM WASTE HEAT" filed April
28, 2009, which is herein incorporated by reference.
FIELD OF THE INVENTION
This invention relates to internal combustion engines with supplemental steam
power obtained from waste combustion heat and to a combination internal
combustion
(LC.) engine and steam engine.
BACKGROUND
Internal combustion engines although highly developed, dependable and relied
upon for almost all road transportation throughout the world generally lose
about 72-
75% of the fuel heating value through radiation, engine coolant and exhaust.
The
measured brake horsepower of a typical six-cylinder spark ignition automobile
was only
21% of the fuel heating value at 72 MPH and only 18% at 43 MPH, Internal
Combustion
Engine Fundamentals, J.B. Heywood, McGraw Hill 1988 pg. 675. Meanwhile,
increasing fuel prices and shortages mount steadily as world supplies dwindle
and air
pollution problems continue. While there have been several attempts to provide
greater
efficiency in an internal combustion engine by recovering energy from waste
heat, prior
proposals have had marked shortcomings. One prior system developed by BMW
International (U.S. patent 6,834,503) requires, in addition to the internal
combustion
CA 02701481 2010-04-27
s
engine, an entirely separate steam expander that is connected to the internal
combustion
engine by a belt to recover power from engine coolant and an exhaust powered
steam
generator. This arrangement adds considerably to the size, weight and expense
of the
power plant as well as placing limitations on thermal recovery. Because of
space
constraints in a vehicle, the volume and weight of the complete unit is
critical. Porsche
AG developed a waste heat turbine that was geared to an I.C. engine (U.S.
patent
4,590,766).
The present invention aims to provide a way to recycle steam continuously in a
closed circuit (no steam exhaust) through a high efficiency expander where
economy of
operation is the prime consideration while the same time improving I.C.
emissions.
Attempts have been made to combine a gas and steam engine for recovering waste
engine heat, examples of which are the Still engine (GB Patent 25,356 of 1910
and
28,472 of 1912 and U.S. Patent 1,324,183) and Mason patent 3,921,404. Still
has a
cylinder cover below the piston that provides a thin annular chamber which
allows steam
to flow in and out between the cover and the piston from an opening in the
cylinder wall.
In a counterflow engine, steam pressure throughout the entire cylinder falls
close to
atmospheric during the entire exhaust stroke producing a drop in steam
temperature
which cools cylinder walls allowing condensation of the steam admitted on the
next
power stroke. This robs the engine of power that would otherwise be available
by
reducing the mean effective cylinder pressure of the incoming charge of steam.
However,
the efficiency of steam engines operating on what is known as the uniflow
principle
achieve much greater efficiency than in a counterflow steam engine by reducing
the
condensation of steam. The inventor of a steam-only uniflow engine described
in U.S.
patents 2,402,699 and 2,943,608 reported tests showing a thermal efficiency of
38.2% at
3450 RPM. A double acting hollow piston uniflow engine is described in Marks
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Standard handbook for Mechanical Engineers, 1987 Section 9-37 as the "last
great
improvement in design" but it is unsuited for use as a combination internal
combustion
and steam car engine in part due to overheating of the piston.
One object of the present invention is to provide a combined internal
combustion
and steam engine that overcomes thermal inefficiencies inherent in prior
combination
engines but has the advantage of utilizing I.C. components (piston, cylinder,
connecting
rod and crankshaft) and efficiency gains that result from sharing some of the
I.C.
mechanical losses as well as having a compact unobstructed combustion chamber
without pockets or extensions as present in an F head (opposing valve) engine
thereby
permitting a high performance, high compression four I.C. valve hemispherical
chamber
construction. A more specific object is to provide a combination engine in
which internal
combustion and steam act on the same piston without steam condensing on the
cylinder
or piston walls or heads upon admission so as to eliminate condensation losses
previously inherent in prior double acting combination engines. To accomplish
this, the
invention must provide an I.C. steam engine with protection against losses
inherent in
filling the clearance space or those due to chilling of steam chamber walls by
low-
pressure exhausted steam as good or better than in what as known as a uniflow
engine.
An important requirement in a double acting I.C. and steam engine is the need
for a
mechanism that uses the least possible added cylinder length to minimize
engine size and
weight. However, it is also necessary to prevent burnt I.C. gas/oil and blow-
by gas from
contaminating the steam and thereby reducing steam generator and condenser
efficiency.
The invention aims to add as little as possible to the cylinder length to
accommodate
steam yet not contaminate the steam with oil or combustion products. Another
general
objective of the present invention is to provide a power source for more
efficiently
utilizing waste heat that is built into the internal combustion engine itself
so that a
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CA 02701481 2010-04-27
separate steam engine or expander is unnecessary, making possible better
recovery of
waste energy from the internal combustion engine as well as a reduction in the
over-all
volume of the power unit and its production cost together with improved
operating
flexibility so that the engine is well adapted for powering vehicles
especially cars, buses,
trucks, locomotives or aircraft. It is a more specific object of the present
invention to
obtain the outstanding efficiency advantages of a combustion piston having an
adjacent
steam chamber that is able to provide both an effective zero steam chamber
clearance
and a gain in mean cycle temperature. Another object is to make possible
reliable steam
admission timing while providing variable steam cutoff in an engine that
derives power
from steam and combustion acting upon a piston yet is flexible enough to
operate
efficiently with large variations in load and steam generator output. Yet
another object is
to more efficiently recover lost combustion heat by conductive transfer to a
working
fluid within the engine itself as well as a more efficient way of recovering
waste heat
from T.C. engine coolant and from engine exhaust gases. Another more specific
object is
to provide a way to capture and remove oil and blow-by gas before it can enter
a steam
line. Still another object is to find a way to accurately vary steam cutoff in
an internal
combustion-steam hybrid engine while being able to recompress residual steam
to
throttle pressure within a combustion piston.
These and other more detailed and specific objects and advantages of the
present
invention will be better understood by reference to the following figures and
detailed
description which illustrate by way of example but a few of the various forms
of the
invention within the scope of the appended claims. Topic headings are for
convenience
of the reader and not intended to be in any way limiting.
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BRIEF DESCRIPTION OF THE DRAWINGS
Fig. 1 is a semi-diagrammatic vertical sectional view of one cylinder of an
engine
in accordance with the invention with the combustion cylinder head elevated
and rotated
90 degrees about a vertical axis to show the valves, the piston being shown at
the top
dead center position;
Fig. 2 is a partial view of the lower half of the cylinder of Fig. 1 on an
enlarged
scale showing the piston at the bottom dead center position;
Fig. 3 is a perspective view of the crosshead and upper portion of the
connecting
rod;
Fig. 4 is a transverse vertical sectional view of the crosshead taken on line
4-4 of
Fig. 3, also showing the crosshead guide column;
Fig. 5 is a greatly enlarged partial vertical sectional view showing a portion
of the
cylinder wall and piston at the bottom dead center position;
Fig. 5A is a view similar to Fig. 5 to illustrate additional oil rings;
Fig. 6 is an exploded perspective view of the steam cylinder head or cap and
piston;
Fig. 7 is a partial vertical cross sectional view of the steam cylinder head
or cap
on an enlarged scale to show the double seated balanced poppet valve;
Fig. 8 is a vertical sectional view showing another form of steam admission
valve;
Fig. 9 is a schematic diagram of the invention as applied to a 4-cylinder car
engine showing steam circulation;
Fig. 9A is a schematic diagram to show the source of supplemental combustion
air fed to the afterburner;
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Fig. 9B is a diagram similar to Fig. 9 showing an alternative condenser and
heat
recovery circuit;
Fig. 10 is a vertical sectional view of a stepped piston engine in accordance
with
the invention;
Fig. 11 is a perspective view of the piston of Fig. 10;
Fig. 12 is a schematic diagram showing the recovery of braking energy;
Fig. 13 is a partial vertical sectional view showing a bump valve in
accordance
with the invention;
Fig. 14 is a schematic diagram showing the control used for changing engine
displacement during operation;
Fig. 15 is a partial vertical sectional view of the center of the cap showing
the
piston rod packing and crosshead;
Fig. 16 is a perspective view of one form of metallic packing;
Fig. 17 is a perspective view of another form of metallic packing;
Fig. 18 is a diagrammatic perspective end view of one cylinder and the steam
camshafts of an engine having two steam admission valves in series within a
combustion
piston according to the invention;
Fig. 19 is a diagrammatic vertical sectional view on line 19-19 of Fig. 18 to
show
the steam camshafts and crankshaft;
Fig. 20 is a partial vertical sectional view of the lower cylinder head or cap
and
cylinder similar to Fig. 19 on a slightly larger scale;
Fig. 21 is a diagram showing the valve opening sequences of a pair of series
connected steam admission valves;
Fig. 22 is a diagrammatic end elevational view partly in section that shows
optional concentric steam admission valves according to the invention;
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Fig. 23 is a diagrammatic perspective view partly in section broken away to
show
the cylinder cap, steam admission valves and the crosshead guide with a pipe
to supply
steam directly to the admission valves;
Fig. 24 is a diagrammatic vertical cross section of a portion of the cylinder
cap
showing an auxiliary clearance volume chamber and a pressure regulating valve;
and
Fig. 25 is a schematic diagram of one form of engine installation assembly and
engine
control.
SUMMARY OF THE INVENTION
This invention concerns a high efficiency composite internal combustion and
steam engine especially suited for use in cars and trucks which includes one
or more
combustion chambers for burning fuel to power a piston by combustion as well
as at
least one expandable chamber within the engine that is powered by steam
generated from
what would have been waste heat from the combustion chamber. Previous systems
for
recovering waste exhaust heat and waste heat from the combustion chamber
coolant in a
dual cycle engine have been inefficient. To overcome this and other
deficiencies, the
present invention provides a combined cycle engine which employs the advantage
of
using high temperature, i.e. superheated steam with a way of accomplishing
uniflow
steam operation inwardly of each internal combustion piston to improve
operating
efficiency as well as benefiting from a way to provide variable steam cutoff
through the
use of one or a pair of series connected, inwardly retractable, steam pressure
balanced
valves that are located in a cylinder cap which is sealed within each piston
operating in
cooperation with steam recompression and a provision for achieving effective
clearance
volume changes that vary with engine speed to thereby further increase
efficiency and
the specific power output from the waste heat energy recovered. In one example
of a
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cutoff control, a camshaft is coupled for changing the phase of one or a pair
of steam
admission valves to vary valve overlap, thus providing continuous regulation
of the
steam cutoff to further reduce specific fuel consumption. Depending upon the
application
of the engine, the cylinder cap which is placed adjacent to a steam exhaust
port can be
unheated but if advantageous is capable of being heated to the temperature of
the
superheated steam supply or, if desired, can have a device for providing an
intermediate
controlled degree of heating to provide optimum power. Engine coolant can be
evaporated in the engine-cooling jacket to form steam which is then
superheated by I.C.
exhaust gases within an engine exhaust manifold for powering the steam
expansion
chamber within each piston. Heating of the cylinder cap makes it possible in
some
engine applications to achieve high efficiencies which surpass those in what
is known as
a uniflow steam engine to provide additional power from waste combustion heat;
an
efficiency level that is much higher than in an ordinary counterflow steam
engine. The
engine also has the flexibility needed under non-uniform steam generator
pressure and
engine load conditions that occur in vehicles through a provision for variable
steam
displacement. Another aspect of the invention concerns a more efficient way to
recover
combustion heat that is contained in the combustion chamber coolant and in the
I.C.
exhaust gas using an exhaust powered superheater comprising an engine exhaust
manifold for supplemental combustion of unburned fuel while also providing for
the
direct conduction of the heat produced in the combustion chamber to increase
the
enthalpy of expanding steam within the steam expansion chamber inside of each
piston.
The invention thus provides an improved heat recovery, heat exchange, steam
generator
and superheater system for generating steam with a way to better construct a
steam
expansion chamber, steam cylinder head and valving, heated steam exhaust area,
a
provision for steam recompression to admission pressure inside of a combustion
piston
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so as to achieve an effective zero clearance and a gain in mean Rankine cycle
temperature along with a steam supply arrangement that is able to act on each
piston
within an I.C. engine so as to more effectively economize on fuel, make a more
efficient
combined gas and steam engine, balance the steam displacement with steam
generator
output to use steam more efficiently, and provide other features that will be
apparent
from the following description.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
Refer now to the drawings. Shown diagrammatically in Figures 1-8 is a
combination internal combustion engine and steam engine 10 that has a cylinder
12
containing a cup shaped trunk style piston 14 which, unlike ordinary pistons,
is machined
and ground to precise tolerances both outside at 16 as well as in the inside
at 18 and is
positioned to reciprocate within an annular space 11 between the inside wall
12a of the
cylinder 12 and a stationary steam cylinder head. While a single cylinder and
piston is
shown for convenience in some views, the invention is of course applicable to
multi-
cylinder engines as well.
The steam cylinder head (Figs. 1-8) which is located within the piston 14
comprises a flat hub, disk or circular cap 20 that may be, say, '/4 to '/2
inch in thickness
supported at the free upper end of an integral crosshead guide column 20a
which is
secured to the crankcase 21 by bolts 21 a. The disk or cap 20 acts as a top or
end cap for
the guide column 20a as well as one end of a steam chamber 44 and has at its
outer edge
a cylindrical surface 19a as a part of a downwardly extending collar 19 that
is
dimensioned to provide a sliding fit within the piston 14 and is grooved to
support
compression rings 20b which provide a sliding seal with the inner cylindrical
surface of
the piston 14. It can be seen that the cap 20 traverses the cylinder 12 in a
location that is
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spaced from its ends. Mounted within the cylinder cap 20 between the outer
surface 19a
and the center of the cap 20 is the steam inlet valve 48. The space between
the column
20a and the cylinder 12 comprises a steam chest 46 containing high-pressure,
high
temperature steam at substantially throttle pressure which is admitted to the
steam
expansion chamber 44 inward of the piston 14 when valve 48 is opened, driving
the
piston 14 upwardly. Cylinder 12 can be independently bolted to the crankcase
62, if
desired, to accommodate steam's thermal disparities rather than being cast en
block as
shown.
At the top of cap 20 is a bushing 22 suitably sealed with packing for a piston
rod
24 which is secured to the piston by a nut 24a that can be notched for a
spanner wrench.
The piston rod slides through the bushing 22 and is secured to a connecting
rod 28 for
transmitting power to a crankshaft 30 inward of the piston. At the inner end
of the piston
rod 24 is a piston style crosshead 25 with a partially cylindrical outer
surface 25a (Fig. 3)
that slides within an inner cylindrical bore 27 of the cap supporting column
20a which
serves as a crosshead guide to take up the side thrust of the connecting rod
28.
Consequently, piston friction is reduced and piston slap common to most I.C.
engines is
eliminated. The crosshead is coupled to the connecting rod by a wrist pin 26.
The alloys
used in the piston 14 and cap 20, are selected to provide a predetermined
balanced
amount of expansion during startup. When an aluminum piston is used, the
interior wall
18 can be electroplated with porous chromium by a well-known method or covered
by a
steel sleeve (not shown) to provide a hard piston ring contact surface. The
steel sleeve
can be spaced slightly from the aluminum piston skirt in most places to reduce
weight
and heat loss from steam within chamber 44. In operation, the skirt of the
piston 14
reciprocates in the annular space 11 between the cylinder wall 12 and the cap
20 as steam
CA 02701481 2010-04-27
is admitted into the steam expansion chamber 44 inside the piston from the
steam chest
46 through the admission valve 48 to raise the piston.
A conventional internal combustion chamber 34 above the upper face 14a of the
piston 14 is enclosed at the top of the cylinder by a cylinder head 35 (shown
90 out of
rotational alignment with the cylinder so that the valves can be seen) which
has an inlet
valve 36, an exhaust valve 38 and port 37, chambers 39 for coolant
circulation, and a
spark plug 40 operating as a four stroke (Otto) cycle I.C. engine that bums
gasoline or
other fuel in the combustion chamber 34 but which can be a diesel engine or a
two stroke
cycle engine, if desired. The combustion chamber 34 is cooled by a coolant at
39 circulated
through a water jacket 12b of the cylinder 12 is compact, unobstructed, has no
side pockets
and, if desired, can even be of high performance, high compression, four
overhead I.C. valve
hemispherical construction to avoid detonation.
Within the wall of the cylinder 12 and extending around it nearly in alignment
but slightly above the top of cap 20 is a steam exhaust manifold 50 which
communicates
with the interior of the cylinder 12 through spaced steam exhaust ports 51. It
can be seen
in Fig. 2 that the steam inlet valve 48 and steam exhaust ports 51 are located
in
approximate lateral alignment but the exhaust ports are outward of the cap 20
at a
slightly higher elevation. In operation, exhaust gas expelled through the
exhaust port 37
of the I.C. engine passes through a steam generator to be described below
which recovers
waste heat by boiling water or other heat transfer liquid to form steam that
is admitted by
the inlet valve 48 into the steam expansion chamber 44 within the piston from
the steam
chest 46 through an inlet passage 46a, 46b (Fig. 7). Exhaust steam escapes
through the
steam exhaust manifold 50 to low-pressure steam return line 52 when the piston
reaches
the top dead center position as exhaust openings 14b in the piston skirt
become aligned
(Fig. 1) with ports 51 to act as an automatic exhaust valve, thus, in effect,
providing a
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self-contained steam engine below each piston 14 of the I.C. engine. It will
be noted that
the automatic exhaust valve opens and closes while the steam chamber 44 is in
an
expanded state. Exhaust steam is condensed, then reheated and continuously
recirculated
back to the steam expansion chamber 44 in a sealed circuit that is separate
from the I.C.
engine intake and exhaust gas thus seldom requiring replacement.
It can be seen that the cap 20 serves as the lower (steam) cylinder head for
the
steam expansion chamber 44, seals the chamber, provides support for the steam
inlet
valve 48 and establishes the clearance volume of steam chamber 44 which is
purposely
kept small to insure efficient operation (see Fig. 2 which shows the piston at
the bottom
dead center position). It is important to note that since both the inside top
wall of the
piston and top wall of the cap have the same shape (a flat plane), the
clearance volume
can be made as small as desired. The arrangement of chamber 44, cap 20, steam
chest 46,
and piston 14 as shown makes it possible for the entire lower end of the steam
expansion
chamber 44 to be steam jacketed including the steam inlet valve 48 and the top
surface of
the cap 20 which may therefore, when desired, be kept close to the elevated
temperature
of the steam chest 46 e.g. 1000 P.S.I. @ 850 F thereby preventing loss of
power due to
chilling or steam condensation on those parts within chamber 44. It will be
noted that the
exhaust ports 51 unlike uniflow ports of an ordinary steam engine are located
on the
cylinder wall adjacent to a heated engine surface, namely, the cylinder cap
20, all of
which except around piston rod 24 can be heated externally throughout
operation when
advantageous by the steam chest 46. Low-pressure steam is exhausted through
line 52
when chamber 44 is over 50% and preferably over 75% expanded to its full
volume. As
illustrated, the exhaust will begin to open when chamber 44 is about 89% fully
expanded. After port 51 closes, throughout substantially the remaining inward
stroke of
the piston, residual low-pressure steam is recompressed to reach admission
pressure. To
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reduce heat loss, the wall of the guide 20a can be made of two concentric
tubes with a
dead air space between them.
The construction shown in Figs. 1-8 produces a marked improvement in
operating efficiency compared to a conventional counterflow engine. For
example,
assuming a 800 p.s.i. throttle pressure and a 10% cutoff, the uniflow steam
rate of the
engine described and shown in Figs. 1-8 is calculated to be 8.2 lb./HP-Hr,
while in an
equivalent counterflow engine the steam rate is calculated to be 11 lb./HP-Hr
so that the
invention is able to make possible a 25% improvement. At a 12.5% cutoff, the
engine of
Figs. 1-8 is calculated to have a steam rate of 8.31b./HP-Hr. vs. 10.4 for a
conventional
counterflow engine (a 20% improvement).
The piston, steam exhaust valve and cap 20 are constructed to enable the
inward
stroke to bring residual steam up to the admission pressure. This produces an
effective
zero clearance in chamber 44 so that the entire steam mass as it enters is
totally
consumed by admission and expansion work and is therefore more efficiently
utilized
while at the same time achieving a gain in mean cycle temperature. Efficiency
can be
better than a uniflow steam engine because unlike the uniflow engine, where
the piston
surface adjacent the exhaust valve cannot be heated, here the entire adjacent
cap 20
(Figs. 1 and 2) above high-pressure steam chest 46 can be externally steam
jacketed and
thus heated continuously when it is advantageous by the steam chest 46 so as
to prevent
chilling the incoming charge of steam. It is of salient importance to note
that the
invention enables low-pressure steam exhausted through line 52 to be kept away
from
the heated area below cap 20.
In the past, fiber shaft packing shown at 22a in Fig. 2 of the kind commonly
used
in steam powered vehicles had to be replaced at regular intervals and became
scorched if
high temperature steam was used. To avoid these problems, to prevent condensed
steam
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from seeping into the crankcase and provide long life, an alternative metallic
packing
175 can be used, if desired, with superheated steam as shown in Figs. 15-17.
Packing 175
is composed of a soft alloy such as babbitt or other metal alloy used in
marine steam
engine packing which needs replacement no more often than the piston rings and
bearings etc. Each metallic shaft packing ring 175 is provided with sloped
outer walls
175a resting on a similarly inclined seat 176a provided within one of two or
more
stacked bronze rings 176 that open upwardly toward the steam expansion chamber
44.
The shaft packing rings 175 can be split for example into four pieces as shown
in Fig. 17
or, if desired, can have a single lap joint as shown in Fig. 16. Each packing
ring 176 is
pressed onto its seat for example by means of compression springs 178 that are
mounted
in pockets on the top of each ring 175. The slope of the seat 176a is chosen
so that the
packing will not become too tight causing excessive power to be lost in
friction. In
operation, the packing rings 175 are able to rise slightly off their seats
against the force
of the springs. Optionally, the inside wall 27 of guide 20a has two opposing
air transfer
channels e.g. longitudinal recesses 179 facing the wrist pin to permit free
air flow around
the crosshead except at the extreme upper end of the guide 20a. The metallic
packing
rings 175 are therefore able to cooperate with the adjacent cylindrical piston
type
crosshead 25 provided in the example shown in Fig. 15 with an optional piston
ring 173
for returning any gas in vapor or liquid back through the packing to the steam
expansion
chamber 44 when pressure builds up below the packing as the crosshead 25
approaches
the top dead center position above recesses 179 to thereby prevent seepage
into the
crankcase. The packing 175 and 176 shown in the figures can be supplemented,
if
desired, with an additional ring of fiber packing (not shown). Lubrication is
supplied to
the packing 175 and to the crosshead 25 through an oil line 177.
Alternatively, if desired,
a more conventional crosshead (Fig. 3) can be used with flat sides to permit
the flow of
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air around it. To eliminate any potential steam condensation in the crankcase
62, the
steam exhaust line 52 can be placed in contact with it for heating the oil in
the sump
enough to drive off moisture.
STEAM VALVE STRUCTURE
It is essential that the steam engine assembly be constructed in a way that
will
enable it to operate at relatively high RPMs since a gas engine needs to
operate with a
piston speed around 3,000 feet per minute. The valves must therefore have a
low mass
and be free from a tendency to float at high RPMs. While any suitable steam
valve can
be used, one form of balanced poppet valve 48 is shown diagrammatically by way
of
example in Fig. 7. The valve 48 is a balanced double-seated poppet valve
having two
valve faces 47a and 47b each resting on its own seat at the top and bottom of
chamber
46a within the cap 20. Chamber 44 communicates through chamber 46a and ports
46b
with the steam chest 46 when valve 48 is opened. Resistance to movement that
would
otherwise exist due to high steam pressure is balanced by valve face 47b which
is
slightly larger than 47a resulting in a small upward force from steam plus the
closing
force of compression spring 63 that is easily overcome by the cam 64 (Fig. 2).
During
operation, when the cam 64 lowers the valve 48, steam will flow from steam
chest 46
into the steam expansion chamber 44 below the piston 14 through the valve
chamber 46a
and ports 46b driving the piston upwardly. If the lower valve seat 47b has an
o.d. of 3/4
inches, its spindle diameter is 1/8 inch, the i.d. of the upper valve 47a is
9/16 inches and
the pressure is 750lb/in.2 then the net upward pressure on the valve would be
about 130
lbs. plus the spring force. The closing force can be further offset, if
desired, by a
balancing plunger (not shown) at the lower end of the valve stem. Assuming the
difference between the i.d. of the lower seat and the o.d. of the upper seat
is 1/64 inch,
CA 02701481 2010-04-27
the resulting downward (opening) force would be 12.4 lb. when the pressure in
44 builds
to 750 p.s.i. and 16.5 lb. at 1000 p.s.i. The valve heads can be rotatably
mounted on their
stems if desired to equalize wear with pitched radial blades as shown in Fig.
7 to impart
rotation. To better enable the valve 48 to rotate freely during operation, it
can be
connected to valve lifter rod 48a, if desired, by a rotary coupling such as a
snap ring or
ball coupling within a sleeve 48b (Fig. 2). Rod 48a opens the valve when
retracted by a
rocker 64a supporting a wheel 64b in contact with a three-dimensional cam
indicated
diagrammatically at 64 that is shifted axially in operation by a servomotor S'
which is
controlled by a central engine management (C.E.M.) computer 305 (Fig. 14) to
optimize
the valve cutoff. The rocker 64a is connected to the crankcase 21 by a pivot
(not shown)
located behind wheel 64b. The seats of valve 48 can be flat or conical and of
such
inclinations that they have a common apex to help assure equality of thermal
deformation. Both the valve and the seats can be cast from the same material
to assure
equal thermal expansion. However, if different metals are used, a known form
of Skinner
self-expanding poppet valve can be used having spring metal packing rings
between
upper and lower telescopically related valve parts to allow about .003 to .004
in.
telescopic action. Compared to sliding valves, the double seated balanced
poppet valves
are especially advantageous since being hollow they have a low mass, require
no
lubrication, have a relatively small lift (travel) for a full port opening and
more adequate
admission at very early cutoff in a variable cutoff operation as here. If
desired, two side-
by-side valves 48 can be connected by a bridge (not shown) for being opened by
a single
lifter rod.
Fig. 8 shows an alternative unbalanced poppet valve 31 which is biased
upwardly
to its seated position against a valve seat 32 above it by spring 31a.
Threaded onto the
lower end of the valve 31 is a sleeve 33 with a boss at its upper end that is
lowered
16
CA 02701481 2010-04-27
M
during operation by a valve lifter viz. a rocker arm 35 having a bifurcated
end with an
opening 35a extending around valve stem 31 for pressing on the boss to open
valve 31. The
stem of valve 31 extends through a valve guide 31b which is threaded at its
outer end into
the base 20d of the cap supporting column 20a. Valve 31 can function as a
relief valve on the
down stroke when pressure in chamber 44 exceeds that in 46.
Alternatively, in place of a poppet valve, a bump valve 182 (Fig. 13) can be
used.
The bump valve 182 is opened during operation by a valve lifter 180 threaded
into the
lower surface of the piston crown 14a in a position aligned with valve 182 so
that when
the piston descends, lifter 180 will open bump valve 182 by separating its
annular valve
face 182a from a seat 184 concentric with an inlet port 185 extending through
the top of
the cylinder head 20 adjacent valve 182. Valve 182 is normally seated by a
spring 186
supported by a spring retainer 188 which is threaded through a bridge 190 that
is integral
with the cylinder head and column 20a. The bump valve 182 provides a fixed
cutoff.
More than one, e.g. four bump valves can be provided in which case the piston
14 should
have four aligned valve lifters 180.
WORKING FLUID
The working fluid used for producing steam can be distilled or deionized water
or
mixtures of water and alcohol or other known binary fluid mixtures such as
ammonia
and water as well as hydrocarbon liquids such as isobutane or isopentane among
others
which is recycled continuously in a sealed circuit. The term "steam" herein is
used
broadly to include vapors of these and other working fluids as well as steam
from water.
To prevent freezing when water is used, a small pilot light which consumes
only a quart
of fuel per week that has been developed by Saab-Scania, Inc., can be used in
cold
weather. A self-draining system has also been previously developed to avoid
freezing.
17
CA 02701481 2010-04-27
S
COLLECTION OF BLOW-BY GAS AND OIL
As best seen in Fig. 5 which shows the piston in the bottom dead center
position,
the piston 14 is provided with a pair of compression rings 54 and a special
oil ring 55
that has a pair of vertically spaced apart strippers with sharp downwardly
directed edges
55a and 55b that carry excess oil downwardly through the cylinder to a
circular
collection channel 56 that has circumferentially spaced ports in the cylinder
wall 12a
above the steam outlet manifold 50. Throughout operation, the collection
channel 56
receives and carries excess oil and blow-by gas to the sump 62 through a
return duct 60
before it can enter the steam exhaust manifold 50. Any oil that enters the
feed water is
removed by filtration or by a settling tank or centrifuge (not shown).
Refer now to Fig. 5a which shows an optional alternate form of oil collection
channel. In this alternative embodiment, the collection channel 56 that
returns blow-by
gas and oil to the sump is a continuous groove in the cylinder wall which is
provided
with a pair of centrally contracting slotted wiper rings 59 that, unlike
ordinary piston
rings, have a working face on their interior surface against which the piston
slides during
operation. If rings 59 are used, the skirt of the piston should be iron or
aluminum
jacketed within a steel sleeve or plated with porous chromium. The rings 59
rest in
contact with one another within the groove 56 and press inwardly on the outer
wall of the
piston. They are pinned in place or to one another with their gaps separated.
Each wiper
ring 59 has several radial passages or slots 59a through which oil and blow-by
gas is
carried to the return duct 60 that leads to the crankcase 62. Consequently,
the small
amount of oil and blow-by combustion gas which is forced past the rings 54 and
55 is
caught by the collection channel 56 by wiper rings 59 and is carried to the
sump so as to
minimize contamination of the steam by combustion gases and waste oil.
Remaining oil
and particulate matter is removed from condensed steam by filtration and an
optional
18
CA 02701481 2010-04-27
settling tank or centrifugal separator (not shown) if needed. If the centrally
contracting
wiper rings 59 are used, they can be cut from cast pipe stock of a suitable
diameter, then
spread outwardly so that the cut ends are separated by a distance of about a
%2 aninch
enabling the ring to be placed on a circular burnisher of the same diameter as
the piston
so that the inner working surface can be correctly shaped while expanded to
provide a
circular inside surface having the same diameter as the o.d. of the piston.
Rings can then
be pinned to one another with gaps opposite each other to prevent leakage,
compressed
and installed at the same time.
At the bottom of the piston are two more compression rings 54 (Fig. 2) on
opposite sides of exhaust ports 14b. All of the piston rings 20b, 54 and 55
are held in
place by pins so that their ends will miss the ports in the wall of the
cylinder and piston
14.
SEALING THE CYLINDER CAP AND EXHAUSTING STEAM
Refer now to Figs. 2 and 6. To maintain steam pressure in the steam expansion
chamber 44, the cap 20 has three compression rings 20b, the bottom one of
which can be
an oil ring that skims off excess oil from the inside of the piston 14 and
returns it through
openings in the cap 20 to the sump 62 through line 20c. During operation, the
piston 14
as already noted has a finished cylindrical e.g. polished inner surface which
slides on the
outer surface of the cap 20. When the exhaust ports are uncovered by the
piston openings
l4b acting as an exhaust valve, any moisture on the top of the cap 20 is swept
out of the
cylinder rather than being left in it to be evaporated again during the
exhaust stroke as is
the case in a counterflow engine. As noted previously, on the down stroke, the
remaining
steam in the cylinder is compressed in the clearance space to the admission
pressure.
Lubrication is provided between the cap 20 and the piston 14 by pressure feed
line 20f
19
CA 02701481 2010-04-27
that carries oil from an oil pump (not shown). A spring loaded relief or
bypass valve (not
shown) can be provided in the cap 20, if desired, to prevent excessive
pressures in
chamber 44 during start up or in case the condenser fails. Any oil or
condensate in the
steam chest can be removed through drain 60a.
From the steam exhaust manifold 50; the low-pressure steam passes through pipe
52 to a condenser and then to a steam generator which will be described below.
From the
steam generator where it is turned to steam, the steam flows through line 49
to the high-
pressure steam chest 46 and into the steam expansion chamber 44 through valve
48
thereby completing an endless circuit as it is continuously recycled
throughout operation.
TIMING STEAM ADMISSION
Figs. 2-8 and 13 illustrate how steam is supplied to the cylinder. The steam
inlet
valve 48 is opened by a valve rocker 64a extending normal to the plane of the
drawing
(the free end of which is shown partly in section in Fig. 2) supporting a
wheel 64b in
contact with a camshaft 61 which is geared to the crankshaft 30 for opening
valve 48
each revolution. The rocker is connected to the crankcase by a pivot pin
located behind
the wheel 64b. Together the camshaft 61, cam 64 and rocker 64a comprise a
valve
retractor for opening the steam inlet valve 48 against the closing force of
steam pressure
in chamber 46 and valve spring 63. If mounted on the same camshaft that is
used for the
I.C. valves, the cams for the steam inlet valves should have two lobes 180
apart so that
the steam valves open each revolution. The camshaft 61 can be gear-driven,
e.g. using
helical gears in a known manner for advancing or retarding the camshaft cam 64
thereby
advancing or retarding the steam cutoff. Alternatively, each cam 64 of
camshaft 61 is a
three-dimensional cam contoured along its length as shown in Figs. 1 and 2 to
provide
for different cutoffs by the action of servomotor S' sliding the camshaft 61
axially (Fig.
CA 02701481 2010-04-27
2) as described above. For example, each cam 64 is shaped to provide, say, a
5% cutoff at
one end and a 50% cutoff at the other. Thus, computer 305 by sliding the
camshaft 61 axially
can select an optimum cutoff to provide the most efficient operation and the
best gas mileage
for a vehicle.
ENGINE MANAGEMENT CONTROL
One example of a controller for adjusting each of the engine operating
regulators
including the inlet valve cutoff cam 64 by servomotor S1, the operation of
steam throttle
T by servomotor S2, and the combustion engine operating controls is a central
engine
management (C.E.M.) computer 305 (Fig. 14) with components of the same general
kind
as those used in engine controls for cars and trucks. The C.E.M. is connected
to one or
more input sensors that indicate the condition of operating variables which
are used in
the present invention for minimizing fuel consumption at each steam generator
pressure,
engine RPM and loading. The monitored steam engine operating variables are
used to
continuously control engine operation by adjusting the steam throttle setting
T, the
variable cutoff cam 64 and other engine operating regulators to maximize heat
recovery
and to minimize both fuel consumption and the discharge of pollutants. The
C.E.M.
computer 305 is also used to provide cylinder compounding and to balance
engine
displacement with steam generator output as will be described more fully below
in
connection with Fig. 14 under the heading STEAM DISPLACEMENT CONTROL.
STEAM ENGINE EFFICIENCY AND ENTHALPY LOSSES
As seen in Fig. 2, the entire top wall of cap 20 above steam chest 46
(typically
about 3 to 6 times the top area of column 20a) keeps the cap 20, valve 48 and
piston
close to the temperature of the incoming steam as noted above. The lower part
of the
21
CA 02701481 2010-04-27
cylinder can be insulated, a small part of which is shown at 12c, as well as
the crosshead
guide 27 as partially shown at 20e to minimize heat losses from the steam
chest 46.
Engines of the type described for example in patents 1,324,183 and 3,921,404
operate on a counterflow principle in which steam at atmospheric pressure is
exhausted
as the piston descends. By contrast, the steam engine assembly of the present
invention,
which is located below the crown of piston 14, is able to surpass even the
performance of
the uniflow steam engine which is noted for its unusually high efficiency. In
a uniflow
engine, steam flows out at the center of the cylinder only during the short
period of time
when the piston uncovers the exhaust ports but the piston ends adjacent the
ring of
exhaust ports are unheated. In the present invention, the alignment of ports
14b in the
piston skirt with exhaust ports 51 in the cylinder wall allow steam to escape
briefly to the
exhaust manifold 50 surrounding the cylinder 12 at the upper end of each
stroke when
the chamber 44 is in an expanded state. When the exhaust ports 51 are
uncovered, any
moisture that collects on the top of the heated cap 20 will be swept
horizontally out of
the cylinder, rather than being left for re-evaporation. On the down stroke,
steam left in
the steam chamber 44 is compressed to the admission pressure but unlike the
uniflow
engine both upper and lower adjacent surfaces of the clearance space 44 are
heated (the
crown of the piston 14 being heated by the I.C. combustion chamber and the
lower end
of the cylinder including the cap 20, valve 48, and piston skirt being heated
by the high-
pressure steam in chamber 46) thereby keeping the cylinder and piston hot. In
this way,
potential chilling of engine surfaces is prevented more effectively than in
the uniflow
engine. Steam recompression as in a uniflow engine avoids the intermittent
cooling that
takes place in a counterflow steam engine while the heating of the cap 20
adjacent the
exhaust ports 51 minimizes condensation of the fresh charge of steam thereon,
thus
maintaining a level of efficiency higher than that of the uniflow engine.
Figs. 2, 7 and 8
22
CA 02701481 2010-04-27
=
show how the cap 20 containing the steam intake valve 48 is steam jacketed to
provide
heat for keeping the steam cylinder head surface of 20a heated adjacent the
ring of
exhaust ports 51 throughout operation.
Steam engine efficiency is also enhanced by the direct conduction of heat from
the burning gas in combustion chamber 34 through the top 14a of the piston to
the steam
under the piston. Of the fuel heating energy that is lost when the fuel is
burned, about 8%
is lost during combustion and about 6% during expansion. Much of this lost
heat is
transferred into the crown and upper part of the piston and in turn to the
steam in
chamber 44. The head of the piston can be maintained at a safe operating
temperature
due to the large volume of steam passing through the chamber below the piston.
To further improve efficiency, and make up for the negative torque on the down
stroke (especially at low speeds), the steam chamber 44 is optionally
connected to an
auxiliary displacement chamber 45 (Fig. 2) in the piston head of, say, 2 in.3
for a steam
cylinder displacement of 40 in.3 through a throttling duct or restriction 47
of about 0.18
inch in diameter to act as a variable auxiliary clearance volume for limiting
down stroke
compression pressure in chamber 44 to throttle pressure at various throttle
settings based
on a known flow throttling principle so as to provide a greater effective
clearance volume
at low RPMs due to the greater flow through the restriction while providing
reduced flow
at high RPMs thereby limiting the charge entering the cylinder. This action
provides the
effect of an earlier cutoff at high RPMs for increasing Rankine engine
efficiency. This
feature and the net positive torque of three out of each four strokes provided
by I.C. and
steam power working together results in a reasonably even running engine. A
negative
torque of about 10-12% of the net positive torque of the other three strokes
will occur
only during the I.C. intake stroke. The invention provides more even torque
than an I.C.
engine for a given number of pistons since the steam below each piston will
provide one
23
CA 02701481 2010-04-27
power stroke per revolution and greater torque at reduced RPMs because steam
engines
do not need high revs to develop their power. A 4-cylinder engine will have
six power
strokes instead of two per revolution including four from steam expansion.
Chamber 45
also lightens the piston and is self-draining if any condensate is present.
Lubrication can be supplied in any suitable manner, e.g. by means of
continuous
pressure feed lines 20f and 57 (Fig. 2) and 177 (Fig. 15) which meter oil to
the cap wall,
the piston and cylinder through the cylinder wall as well as by conventional
pressure
feed through the crankshaft and connecting rod or by a splash or oil jet from
the sump
(not shown).
THERMAL ENERGY RECOVERY
Refer now to Fig. 9 which illustrates one preferred form of thermal recovery
assembly and method in accordance with the present invention. In a typical
I.C. engine,
the brake power output is about 25-28% and losses amount to about 72-75% of
the
heating value of the fuel under the conditions of use (see I.C. Engine
Fundamentals by
J.B. Heywood, 1988 page 674). I.C. energy losses include those from the
exhaust, the
coolant and miscellaneous losses such as radiation from the engine and exhaust
system
and exhaust kinetic losses. The system for recovering this thermal energy
shown in Fig.
9 includes the engine 10, a steam generator 100, an economizer 102, a steam
superheater
104, two countercurrent liquid-to-liquid heat exchangers 106 and 108, a
condenser 112
and radiator 110. As outlined above, the low-pressure steam is collected in a
chamber or
manifold 50 surrounding each of the cylinders and is exhausted through the low-
pressure
steam return line 52. The steam return line 52 is connected by line 114 to the
countercurrent flow heat exchanger or regenerator 106 which can consist of a
pair of
tubes mounted one inside the other, the inner one being formed from a material
such as
24
s6_.
CA 02701481 2010-04-27
copper that is a good conductor of heat. Regenerator 106 acts as a secondary
preheater
for preheating the feed water flowing through the inner tube 116 which is then
carried to
the economizer 102 through line 118. The economizer 102 can be in steam
generator
casing 124 but if separate (as shown) better counterflow heat transfer is
assured. The
economizer 102 in the example shown is a countercurrent flow heat exchanger in
which
preheated feed water is heated further by engine exhaust gases that flow in
heat
conductive relationship with the feed water passing through a coil of tubing
120 within
the economizer for providing preheated water to the steam generator 100.
While various steam generators can be used, a monotube steam generator or
flash
steam generator of seamless steel or nickel alloy tubing typically 7/16"
diameter coiled
in a series of flat spirals or wound to form concentric frustocones or as
described in
patent 5,845,609 is preferred. The coils of tubing 122, in the steam
generator, that
receive the preheated feed water from the economizer 102 are in this way
exposed during
operation to the continuous circular flow of hot engine exhaust gas that
enters the casing
124 of the steam generator 100 through tangential inlet opening 126 and leaves
through a
tangential outlet opening 128 in the opposite direction from the flow of feed
water and
steam through the coil 122 to again provide a countercurrent-flow exchange of
heat so
that the last of steam to leave the top coil 122 of the steam generator has
been heated the
most and is exposed to the highest temperature engine exhaust entering through
the
tangential inlet opening 126 at the hottest part of the steam generator 100.
While only a
few coils have been shown in the steam generator, many more are used in
practice. A
length of 5/8 inch o.d. mild seamless steel tubing that makes a total of 24
flat coils of
17.5 feet each totaling 420 feet of tubing (which amounts to 75 square feet of
heating
surface) can provide a 41 HP steam generator that does not cause excessive
I.C. exhaust
back-pressure losses. Although the steam generator shown and described is
preferred,
CA 02701481 2010-04-27
other known water tube steam generators can provide efficient energy recovery.
For
example, a LaMont flash tube steam generator developed an efficiency of 85%
(see Heat
Engines by R.H. Grundy, 1952 pages 452 & 453) and a Benson steam generator
during
tests yielded a thermal efficiency of about 90% (see Theory and Practice of
Heat Engines
by D.A. Wrangham 1960 pages 710 and 711). The present invention is able to
provide
equal or better efficiencies since the temperature of gas in the steam
generator 100 is not
reduced by dilution with a relatively cool incoming air stream as is the case
in an
ordinary steam generator in which air is blown by a fan into the burner of the
steam
generator. The heating rate in a tube type steam generator is also faster than
a standard
steam generator. Some flash tube steam generators can get high steam pressure
in as little
as 15-30 seconds which indicates that quite efficient heat transfer is being
achieved. A
circulation pump in a parallel circuit (not shown) can be provided to maximize
heat
transfer.
Refer now to Fig. 9B which illustrates a somewhat different circuit for steam
and
coolant circulation. In Fig. 9B the steam generator, the auxiliary combustor
and
economizer are the same as already described in connection with Fig. 9 but the
following
components are arranged differently: condensate preheater 160, condenser 162,
filter
164, condensate pump 166 and a thermostatically controlled bypass valve 170.
In
operation, wet steam and condensate from the regenerator 106 flows through the
bypass
valve 170 to the top of condenser 162. However, at lower condensate
temperatures, valve
170 allows a greater fraction to bypass the condenser 162 through pipe 172 to
filter 164
from which it is fed by condensate pump 166 to the first stage countercurrent
flow
condensate heat exchanger 160 where it is heated as closely as possible to a
temperature
approaching that of the hot coolant leaving the engine for a maximum
temperature gain
e.g. to about 100 C-115 C. Engine coolant itself is fed in the opposite
direction by
26
CA 02701481 2010-04-27
w
water pump 109 so that upon exiting through pipe 111, it will have transferred
virtually
all of its thermal load to the condensate while being cooled as it reenters
the engine 10 to
the temperature of the condensate leaving condenser 162. The bypass valve 170
diverts
as little as possible through the condenser 162 for maintaining the condensate
and
coolant at the optimum design temperature selected for the particular engine
10 that is
used. Circulation of the steam to and from the engine through the economizer
102 and
the secondary condensate preheater 106 is the same as already described in
connection
with Fig. 9.
The thermal energy recovery method and apparatus of the present invention, as
described above and particularly in connection with Figs. 9, 9A, and 9B, is
outstanding
for application in an automobile or other vehicle because of its high
efficiency in
recovering waste heat and its relatively light weight. However, other heat
exchange
arrangements are well known as described for example in U.S. Patents
4,087,974;
4,201,058; 4,300,353; 4,351,155; 4,803,958; 6,834,503; 6,964,168; 7,013,644,
some of
the features of which can also be used in recovering waste energy from the
internal
combustion section of the engine of Figs. 1-8. All references cited are
incorporated
herein by reference as fully and completely as though they were reproduced in
full in the
text of the present application.
Figs. 9 - 9B are intended to diagram the circulation of steam and coolant in a
general way. Other well known components (not shown) including temperature and
pressure sensors, check valves, liquid storage tanks, thermostatic engine
coolant valves,
fans for radiator/condenser, pressure gauges and relief valves, drains, and
the like, all
familiar to those skilled in the art are used conventionally. To minimize heat
loss,
components that are at elevated temperatures are thermally insulated
conventionally for
example as shown partially in 12c and 20e in Fig. 2.
27
CA 02701481 2010-04-27
SUPERHEATER ASSEMBLY
In accordance with the present invention, a superheater 104 is provided at the
location of the exhaust manifold of a standard I.C engine. The superheater 104
which is
somewhat larger than a standard exhaust manifold of an ordinary I.C. engine
acts as an
afterburner that forms part of an exhaust manifold for recovering additional
waste energy
while removing some pollutants e.g. CO and hydrocarbons. Inside is a series of
coils 130
of stainless steel tubing for superheating the steam produced in the steam
generator 100
by heat transferred from the engine exhaust gases introduced into the
superheater 104
through exhaust gas inlet pipes 141-144 which are themselves connected
directly to the
exhaust passages 37 in the cylinder head 35. Because the superheater 104 is
between the
steam generator and the cylinders and is connected in close proximity e.g. 2-
10 inches
from the exhaust ports 37 by inlet pipes 141-144, the coils of tubing 130
inside it are
exposed to the greatest heat with steam flowing counter to the flow of exhaust
gases. To
maximize exhaust gas temperatures while also reducing pollutants, heated
secondary air
is injected into pipes 141-144 via injectors supplied with air from a blower
148 via air
supply line 146. Fig. 9A shows how supplemental air entering pipe 147 is
heated by I.C.
exhaust pipe 103 as blower 148 transfers it to I.C. exhaust passages 141-144.
It can
therefore be seen that the coils 130 are exposed to both combustion products;
those
produced in the engine cylinder as well as those that result from the
combustion of
unburned gas that takes place within the superheater due to the injection of
secondary air.
The blower e.g. a positive displacement vane or roots blower 148 can be driven
from the
engine, by an electric motor 150 or by a small capacity exhaust gas or steam
turbine (not
shown) connected to line 114. Exhaust gas entering the superheater 104 through
the
exhaust passages 141-144 can be as high as about 900 C but the most common
range is
about 400 C - 600 C. The auxiliary air supply introduced through the
supplemental air
28
CA 02701481 2010-04-27
supply line 146 will oxidize much of the unburned hydrocarbons and carbon
monoxide
present in the exhaust gas which may amount to as much as 9% of the heating
value of
the fuel. To optimize combustion and increase residence time, the superheater
104 is
made much larger than a standard exhaust manifold, typically around 6-8 inches
or more
in diameter for a four-cylinder engine. Optional swirl guides 105 with pitched
radial
blades give the gas a swirling action and increase residence time within the
superheater
104 to enhance the combustion of unburned gas which is advantageous since it
has been
found that a 1.5% CO removal results in a 220 K temperature rise (Heywood Id.
page
658). It will be seen that the superheater 104 is an afterburner that is made
an integral
part of the exhaust manifold itself where the I.C. exhaust gas at the highest
temperature
enters at several e.g. 4 points with combustion taking place therein where the
monotube
steam generator steam runs in a counterflow direction to incoming exhaust gas
to thereby
provide superheat at the highest temperature since the monotube steam
generator line
passes through the afterburner, entering furthest from the engine and leaving
near the
upstream end of the afterburner. It will also be noted that the steam flows
from the steam
generator into the superheater which receives upstream exhaust gases just as
they exit the
engine and while they are being further heated by the combustion of previously
unburned
hydrocarbons and other combustible gases resulting from the injection of hot
air from the
secondary air supply line 146. Consequently, the invention makes possible the
recovery
of heat from unburned gas and fuel which in an ordinary engine amounts to
about 3-9%
of the heating value of the fuel.
Briefly, the circulation of steam and condensate is as follows: from the
engine
exhaust through lines 52 and 114 then, through regenerator 106 and line 107 to
the
condenser 112. Condensate from the condenser passes through filter 113, pump
115,
through the first preheater 108 then through regenerator 106 (leaving at 220
C - 240 C)
29
CA 02701481 2010-04-27
to the economizer 102, steam generator 124, superheater 104 then through steam
throttle
T, high-pressure steam line 49 and solenoid operated selector valves V to the
steam
chambers 46 and inlet valves 48. The condensed steam is fed by condensate pump
115
from a filter 113 to a primary heat exchanger 108 where it is heated by the
engine
coolant fed in the opposite direction through a line 117 by a pump 109. Thus,
the engine
coolant transfers its heat to the condensate first and is then cooled further
by passing
through the radiator 110 from which it is circulated back through the engine
cooling
passages 35 and 12b via hose 111.
The driver's foot throttle lever can control the throttle set point as a
command to
C.E.M. computer 305 which in turn positions the servos S' and S2 for the
cutoff and
steam throttle respectively to optimize efficiency and maximize gas mileage
continuously while the vehicle is in operation. Engine power can be thus
controlled by a
combination of steam throttle T and cutoff settings to minimize fuel
consumption.
Alternatively, the steam throttle T can be operated for example as described
in patent
4,300,353 with pressure and temperature controlled as described in the Carter
patents
3,908,686 and 4,023,537. The drivers foot throttle can be set to open the
steam throttle T
before the I.C. throttle is opened to use stored energy, if any, before
burning more fuel,
with continued movement of the foot pedal then set to open the I.C. throttle.
STEAM DISPLACEMENT CONTROL
One objective of the invention is to operate efficiently at a variable load
and
steam generator pressure that results from variations in waste I.C. heat and
vehicle
driving requirements. Some steam generator output accommodation can be
accomplished
with the variable cutoff 64. However, the invention also includes a mechanism
for
changing steam displacement to provide the required flexibility under varying
operating
CA 02701481 2010-04-27
conditions. In accordance with the invention, valves V are opened sequentially
by the
C.E.M. computer to increase displacement and are closed sequentially by the
C.E.M.
computer to as closely as possible match steam engine displacement with the
steam
output entropy from the steam generator and superheater, the objective being
to maintain
a more constant ratio between them throughout operation as the engine speed
changes
during startup and under varying traffic conditions. To accomplish this, an
engine load
sensor (torque meter) or a steam pressure sensor at the steam generator or
cylinder is
connected to the C.E.M. computer 305 which opens valves V in a step-wise
manner as
steam generator output rises. Consequently, when the steam generator output is
low, only
one or two steam cylinders are used with more added as the steam generator
output
increases to maintain good efficiency as the heat output of the I.C. engine
changes. For
example, in a steam powered car, other things being equal, lowering steam
engine
displacement from an excessive displacement of 200 CID to 140 CID at a
constant steam
rate of 8.51b./HP-Hr. is calculated to improve fuel economy from 9.9 MPG
initially to
14.4 MPG (a substantial 45% improvement in fuel economy) by better matching
steam
generator output and engine displacement.
One example of an engine displacement control in accordance with the invention
is shown in Fig. 14. It will be seen that steam is fed to the cylinders of the
multicylinder
engine 10 from steam generator 100 through throttle T via steam line 49 and
four
solenoid operated valves V to the engine cylinders as described above. The
steam
generator is provided with sensors for operating variables such as temperature
and
pressure sensors (not shown) which are connected to the engine management
computer
305 by conductors 300, 302 and by conductor 303 for indicating heat supplied
to the
steam generator (Btu/hr). The load on the engine 10 and engine RPM is
transmitted to
the C.E.M. 305 via lines 306 and 308. While the engine is running, the C.E.M.
carries
31
CA 02701481 2010-04-27
out real time computation of the optimum engine displacement under existing
operating
conditions for achieving the best gas mileage. This result is then used to
operate the
solenoid-controlled valves V via conductors 304'-307'. If each steam cylinder
of the
engine 10 has the same displacement, then the valves V are opened in sequence
to
provide engine flexibility with respect to steam generator output by matching
displacement with the operating conditions being experienced so as to increase
the
displacement, e.g. in a step-wise manner from 15 in.3 to 30, 45, and finally
60 in.3 as the
steam pressure, temperature or heat supplied to the steam generator increases
while
reducing the displacement responsive to decreased steam generator output or
engine
load, all according to predetermined operating parameters for maintaining the
best steam
engine efficiency under the conditions of operation sensed by the computer
while the
engine is running as well as before starting or when coasting down. When there
is no
load on the engine, e.g. when coasting, the throttle T or all valves V are
shut by the
C.E.M. computer 305 but this is overridden upon depressing the foot throttle.
Other
operating variables such as I.C. manifold pressure and vehicle speed can be
monitored to
provide additional operating conditions to the controller 305 as well as those
indicated.
In designing the engine, the optimum displacement chosen for steam operation
can be
achieved by selecting a particular piston i.d. or piston rod o.d. If a thicker
piston skirt is
needed, an optional dead air space can be provided in the piston between the
piston skirt
and a concentric liner sleeve (not shown) if desired. Admission valves of
steam
chambers which are receiving no steam can be held open as exhaust valves are
held open
in patent 2,196,980 if desired. To further reduce piston weight and enhance
cooling the
piston crown can enclose a hollow annular chamber (not shown) containing a
quantity of
sodium or other heat conductor to promote heat transfer. See Ford patent
6,904,876 for a
sodium cooled piston and Toyota patent 4,712,600.
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CA 02701481 2010-04-27
STEAM CYLINDER COMPOUNDING
It is well known that a compound steam engine can provide greater efficiency
at
pressures over about 150 pounds per square inch. Using the solenoid operated
valves
(not shown) between the exhaust line 52 of a cylinder and the inlet line 49 of
one or
more cylinders, exhaust steam from one cylinder can be sent to a cylinder made
with a
greater displacement or to two or more cylinders of the same displacement to
achieve
compounding through automatic control thereof by the C.E.M. computer e.g. from
cylinder I
to both of cylinder 2 and 4 with no receiver.
HEAT TRANSFER AND OPERATING TEMPERATURES
As in any engine, it is necessary to maintain components at temperatures that
will
not impair proper lubrication. In a typical water-cooled four stroke spark
ignition engine
that was tested, temperature measurements taken at the top of the piston and
at various
locations around and behind the top ring varied from about 290 C to about 340
C.
Therefore, running the engine 10 with steam at 300 C - 350 C should be
acceptable for
all applications since both aluminum and cast iron pistons can operate with
the piston
head at temperatures ranging from 200 -400 C in a standard I.C. engine. The
exhaust
valve and spark plug can safely run at around 310 C to 340 C. In the present
invention,
the I.C. head and cylinder are both water-cooled. Since the cylinder walls of
the
combustion chamber are cooled by the water jacket, the piston head has the
greatest
potential for overheating. Heat transfer and cooling of the piston 14 is an
important
consideration. As seen in Fig. 2, when the steam inlet valve 48 opens, the
pressurized
steam will stream upwardly and rush across the lower surface of the piston
head 14a and
will thereby carry away excess heat as its enthalpy rises. At typical highway
cruising
speeds as much as about 0.3 lbs. of steam per second will flow across the
lower surface
33
CA 02701481 2010-04-27
of the piston head for carrying away heat so as to avoid overheating of the
piston head
even though the initial temperature of burning gas in the combustion chamber
ranges
from 2000 C-2400 C. It will be noted that the invention makes it possible to
conserve
energy due to the direct conduction of heat from the combustion chamber to the
steam
streaming across the lower surface of the piston head as the steam is
introduced beneath
the piston thereby raising the pressure of the steam while at the same time
combustion
heat also prevents condensation of the steam on and around the piston head. If
desired,
the lower surface of the piston head can be provided with cooling fins to
promote heat
transfer to the steam.
THERMAL LOSSES THAT ARE AVAILABLE FOR RECOVERY
The following tables showing fuel energy utilization and loss are derived from
Heywood Id. 1988, pages 674 and 675. To evaluate efficiency, brake power is
compared
with heat transferred to the coolant and to the exhaust as well as other
losses that
together make up the lower heating value of the fuel being used.
TABLE I
TYPICAL FUEL ENERGY DISTRIBUTION IN SPARK IGNITION ENGINES
Brake Power 26%
Lost to Coolant 23% Available for Recovery
Miscellaneous Losses 8%
Sensible exhaust enthalpy 26% Available for Recovery
Kinetic exhaust enthalpy 3%
Lost to exhaust systems by radiation 5%
Incomplete combustion 5-9% Available for Recovery
100%
Total available for recovery: 54-58%
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CA 02701481 2010-04-27
TABLE 2
FUEL ENERGY DISTRIBUTION OBSERVED IN 6 CYLINDER
SPARK IGNITION AUTOMOBILE
43 MPH 72 MPH
1100 RPM 1800 RPM
Brake Power 18% 21%
Coolant Load 54% 43% Available for Recovery
Exhaust enthalpy 21% 27% Available for Recovery
Misc. radiation from exhaust 2.5% 3.2%
Incomplete combustion 4.5% 5.8% Available for Recovery
100% 100%
Total available for recovery: 79% 76%
The heat rejection rate to the coolant changes under different operating
conditions. In an automobile at low speeds and loads (Table 2), the coolant
heat transfer
rate is much greater, amounting to as much as 2-3 times the brake power.
DETERMINATION OF ENERGY RECOVERY
For a four cylinder spark ignition four stroke automobile engine operating as
set
forth in Table 1 and assuming an I.C. engine rated at 100 I.C. HP, the brake
I.C. power
produced under highway conditions would be 26 HP with 54% of the fuel heating
value
being available for recovery. Assuming the steam generator is 85% efficient
and the
steam engine operates as a high-pressure superheated condensing engine having
an
actual brake power efficiency of 24% (see Heat Engines, Allen Bradley, page
407), the
combined efficiency is therefore 20% x 54 HP or 10.8 HP for an improvement on
the
order of 41.5%. Using the lost energy percentage values of Table 2 derived
from a
vehicle traveling at either 43 or 72 MPH, the potential recovery is 20% x 79%
= 15.8%
which amounts to an 87% improvement in gas mileage at 43 MPH. At 72 MPH with
the
energy available for recovery of 76% x 20% efficiency = 15.2% of the thermal
energy
available for a brake power improvement of 72%. Consequently, the improvement
in gas
CA 02701481 2010-04-27
4 S
mileage that can be achieved through the use of the invention based on
empirical test
data varies from about 41 % to about 87% above the gas mileage otherwise
achieved.
RECOVERY OF UNBURNED FUEL ENERGY
As already noted, unburned fuel contributes about 3-9% to the waste exhaust
gas
enthalpy. Burning it can raise the temperature of the exhaust gas as much as
200 K.
Running at high speed, engine exhaust from a typical vehicle engine will range
from
about 400 C-900 C.
As seen in Fig. 9A, an inlet pipe 147 connected to the inlet of the air pump
148 is
in a countercurrent heat exchange relationship with the exhaust pipe 103 for
recovering
exhaust heat and transferring it to the air that is forced by the pump 148
through the air
injector line 146 into each of the exhaust passages 141-144 to supply oxygen
for burning
hydrocarbons within the superheater 104. The exhaust passages 141-144 enter
the
superheater 104 through tangential ports 145 which cause the exhaust gases to
swirl as
they flow through the superheater 104. The swirling action is sustained by the
pitched
swirl plates 105 which act like a helix to help ensure complete mixing of the
hot
secondary air so as to achieve total combustion as well as reducing the
emission of
pollutants. Thus, the tangential inlets 145 begin the swirling motion and
swirl plates 105
promote both swirling and agitation so as to assure complete mixing of the
secondary air
which cooperates with the large (6" to 8" diameter) size of the superheater to
oxidize the
unburned constituents thereby raising the temperature of the exhaust to
between 600 C -
900 C or more. This temperature is well above that required for virtually
complete
combustion of both unburned hydrocarbons (requiring about 600 C for 50 ms.)
and CO
(requiring about 700 C for about 100-150 ms.). By placing the steam coils 130
within it,
the superheater 104 serves as a burner or furnace due to the injection of hot
secondary air
through air inlet line 146 causing the steam in the last coils 130 to reach a
higher
36
CA 02701481 2010-04-27
temperature than it would in an ordinary engine exhaust manifold thereby
further
improving Rankine engine efficiency in the engine expansion chamber 44 due to
the
higher temperatures and pressures achieved.
Thus, running at a speed of about 70-75 MPH with the exhaust gas in the
superheater at a temperature of around 900 C - 1000 C due to the additional
heat
provided by combustion of unburned fuel, steam at the throttle T at designed
flow rates
can be heated to over 500 C (932 F). It is apparent, therefore, that high
steam
temperatures and pressures necessary for efficient Rankine engine operation
are readily
obtained by means of the invention.
RECOVERING MOMENTUM OF VEHICLE
Refer now to Fig. 12 which illustrates the recuperation of kinetic energy
during
braking in accordance with the invention. A four cylinder automobile engine
150 of this
invention e.g. Figs. 9-9B is shown connected by a drive shaft 151 to a drive
wheel 152
resting on the ground. The illustration shows a solenoid operated selector
valve 153 in
steam outlet line 52 and a solenoid valve 158 in the steam line leading to the
steam
generator. Also in line 49 is a solenoid selector valve 159 that is
connectable via a first
airline to a check valve 154 for receiving ambient air and a check valve 155
coupled to
compressed air tank 156. The selector valve 159 can also connect line 49 via a
second
airline through check valve 157 to the compressed air tank 156. When the
brakes are
applied, the poppet valves 48 inside the engine 150 are held open by the
engine
management computer and selector valve 159 connects tank 156 to line 49 via
valves
154 and 155 so that the changing volume of the steam expansion chambers 44
(Figs. 1
and 2) forces air under pressure into tank 156 until the vehicle stops thereby
storing
kinetic energy in the form of compressed air. When the accelerator is
depressed by the
37
CA 02701481 2010-04-27
driver, the solenoid operated selector valve 159 connects tank 156 to line 49
through the
check valve 157 while valve 48 operates normally as it does during steam
operation
thereby propelling the vehicle ahead due to the expansion within the steam
chamber 44
of the compressed air that was stored during braking. Compressed air is vented
through
valve 153. In this way, the invention provides a vehicle that is propelled by
a
combination I.C. and steam power that was derived from thermal energy
recovered from
waste I.C. engine heat and momentum by selectively connecting a steam
expansion
chamber of the engine to a compressed air storage tank through valving that
directs the
flow of air from the expansion chamber either into a storage tank during
braking of the
vehicle or to move ahead, connects the steam expansion chamber to the air tank
for
directing compressed air from the air tank into the steam expansion chamber.
With a cold
engine and no steam pressure, the compressed air is able to provide a fast
start even
before raising steam as well as enhanced fuel efficiency during start up. It
can therefore
be seen that the invention provides an energy storage unit arranged such that
the brake is
operatively connected to the wheels for transferring momentum from the wheels
152 of
the vehicle during braking (when engine power is not applied to the wheels) to
the
energy storage unit such as the compressed air tank 156 or any other energy
storage device
for later application to the wheels for moving the vehicle forward.
STEPPED PISTON ALTERNATIVE
Refer now to Figures 10 and 11 which illustrate an alternative embodiment of
the
invention that employs a stepped piston 80. While this variation of the
invention makes
use of a different piston construction, it is also able to recover power from
waste I.C.
heat. The modified engine 79 shown in Figures 10 and 11 operates to achieve
the same
general objectives already outlined; better efficiency and gas mileage, but
instead of
38
CA 02701481 2010-04-27
r
having a steam expansion chamber inside the piston, the stepped piston 80 is
provided
with compression rings 80a at its upper end and an enlarged diameter 80a in
the skirt
area at its lower end that is provided with compression rings 80b. The
combustion
chamber 34 as well as the cylinder head 35 and its components are the same as
already
described above. The piston 80 fits sealingly for sliding motion in a cylinder
82 that has
an enlarged cylindrical lower bore section 84 to accommodate the enlarged
skirt portion
80a, thus providing an annular steam expansion chamber 86 in the enlarged bore
section
84. Both its combustion chamber 34 and the steam chamber 86 are adjacent the
piston
and chamber 34 is located adjacent the bottom portion of the piston 80.
Lubricating oil is
supplied under pressure through lines 97 to the inside wall of the cylinder.
In operation,
steam enters the steam expansion chamber 86 through an inlet valve 92 and port
91 from
high-pressure valve chest 94. When the piston reaches bottom dead center and
uncovers
openings in the cylinder, steam is exhausted through a steam manifold 88
provided in the
part of the engine cylinder block 89 that encircles the enlarged bore 84 from
which it
flows through a closed circuit like that described above including a condenser
(not
shown) through a check valve e.g. reed valve 99 via a low-pressure steam
return line 90.
The piston 80 located as shown between the combustion chamber 34 above the
piston
and the steam chamber 86 is connected to a crankshaft 87 by a wrist pin 83 and
connecting rod 85. Waste combustion heat from the I.C. chamber 34 is recovered
in a
steam generator and steam circuit as shown and described above in connection
with Figs.
9-9B to provide steam to chamber 86 for driving the piston 80 downwardly each
revolution of the crankshaft 87. This is not a preferred embodiment in part
because the
large surface to volume ratio of the annular steam expansion chamber 86 will
promote
cooling resulting in reduced steam pressure. Flow restriction around the
piston will also
39
CA 02701481 2010-04-27
produce greater breathing losses, longer rings 80b result in increased
mechanical losses and
there is no heated surface adjacent steam outlet 88.
DESCRIPTION OF FIGS. 18 TO 25
Refer now to drawing Figures 18 through 25 wherein the same numerals refer to
corresponding parts described above. The engine 10 has one or more cylinders
12 each
with a combustion piston 14 slideably mounted therein and sealed by piston
compression
rings 54, 56, a combustion chamber 34 outward of the piston 14 and a steam
expansion
chamber 44 in the cylinder 12 inward of the piston, i.e. between the piston
and the
cylinder cap 20. In general, the engine shown is constructed and operates as
described
above in connection with Figs. 1-8. Differences in construction and operation
will now
be described. Mounted within the cylinder cap 20 to control the flow of steam
from the
steam chest 46 through the cap to the steam expansion chamber 44 above its
upper
surface are two series connected steam admission valves 200 and 202 which it
will be
noted are enclosed within the piston 14. The space between the crosshead guide
20a
which supports cap 20 and the cylinder 12 defines the steam chest 46 as
described above
which contains high-pressure, high temperature steam at substantially throttle
pressure
that is admitted to the steam expansion chamber 44 below the piston 14 when
both of the
valves 200 and 202 are in the open position thereby driving the piston 14
outwardly.
Valve control for regulating the supply of steam to the expansion chamber 44
during
each expansion stroke is accomplished in this example by changing the phase of
one of
valves 200 or 202 as will be described more fully below. Valve 200 is ported
through the
cylinder cap 20 about midway between the piston rod 24 and the cylinder 12
(Fig. 18).
During operation, the skirt of the piston 14 reciprocates as already described
in the
annular space 11 (Fig. 20) between the cylinder wall 12 and the cap 20.
CA 02701481 2010-04-27
In the manner described in connection with Figs. 1-8, a conventional internal
combustion chamber 34 above the upper face 14a of the piston 14 is enclosed at
the top
of the cylinder by a cylinder head 35 which has an inlet valve 36, an exhaust
valve 38, a
steam jacket 12b, 39 (shown in Figs. 1 and 2) for coolant circulation between
the
cylinder and the head with a spark plug 40 operating as a four stroke (Otto)
cycle or two-
stroke I.C. engine that burns gasoline or other fuel in the combustion chamber
34 all as
described above. The internal combustion section of the engine can run as a
spark or
compression ignition engine or combination of both wherein spark ignition is
used for
starting the latter. The combustion chamber 34 is cooled by a coolant
circulated through
the cooling jacket 12b (Figs. 1 and 2) and then through 39 (Fig. 1).
Within the wall of the cylinder 12 and extending around it nearly in alignment
but slightly above the top of cap 20 is the steam exhaust manifold 51 (Fig.
19) which
communicates with the interior of the cylinder 12 through circumferentially
spaced
steam exhaust ports 50 as described in more detail concerning Figs. 1 and 2.
It can be
seen in Fig. 19 that the steam inlet valves 200 and 202 and steam exhaust
ports 50 are
located in approximate lateral alignment but the exhaust ports are at a
slightly higher
elevation outward of the cap 20. It is also important to notice that by
exhausting through
the cylinder wall 12, the invention makes it possible to keep low-pressure
exhaust steam
away from the heated area below cap 20 where it could reduce the enthalpy of
steam in
the steam chest and in the incoming steam. The steam cylinder head or cap 20
has a
peripheral collar 19 that is slideably and sealingly engaged with the inside
wall of the
piston. The steam exhaust valve also includes one or more ports 14b in the
skirt of the
piston 14 that are each spaced from the free lower edge of the piston by a
skirt portion 18
(Fig. 19) which is large enough to overlap the collar 19 of the cylinder cap
and maintain
a seal when the piston is at the top center position shown in Fig. 1 as well
as in
41
CA 02701481 2010-04-27
application S.N. 12/075,042 filed March 7, 2008, now patent which
is incorporated herein by reference.
As in Figs. 1-8, the cap 20 of Figs. 18-23 serves as a lower steam cylinder
head
for the steam expansion chamber 44, seals the chamber, and in this embodiment
provides
support within piston 14 for both of the two series related steam inlet valves
200 and 202
as well as establishing the clearance volume of steam chamber 44 which is
purposely
kept small to insure efficient operation (see Figs. 2 and 19).
In one form of the invention, the entire lower end of the steam expansion
chamber 44 is steam jacketed including the steam inlet valves 200 and 202 and
the top
surface of the cap 20 which are therefore kept close to the elevated
temperature of the
steam chest 46 e.g. @ 850 F thereby preventing loss of power due to steam
condensation
on those parts within chamber 44. It will be noted that the steam exhaust
ports 50 unlike
uniflow steam exhaust ports of a conventional uniflow steam engine are in this
case
located on the cylinder wall adjacent to an engine surface that can be heated,
namely, the
fixed cylinder cap 20, all of which except the center part close to piston rod
24, can be
heated externally, if desired, throughout operation by the high pressure steam
in the
steam chest 46. Steam jacketing of the cylinder end cap 20 in accordance with
the
invention is beneficial in keeping the inner wall of the piston hot and can
also be
beneficial especially in small bore engines and under reduced loads with
earlier cutoffs
or at lower-range steam temperatures and pressures, e.g. during idling as well
as start and
stop driving when steam is wetter. Low-pressure steam is exhausted through
line 52 as
before. As illustrated, the exhaust valve will begin to open and will close
fully when
chamber 44 is about 89% fully expanded and is fully open at the top center
position.
Consequently, steam recompression will begin on the down stroke following
steam
42
CA 02701481 2010-04-27
exhaust as the piston 14 leaves the upper end of its stroke, i.e. the top dead
center
position.
The engine of Figs. 18-22 produce a marked improvement in Rankine efficiency
compared to a conventional counterflow steam engine in which steam enters and
leaves
through the same valve. When port 51 closes, steam is recompressed
substantially
throughout the remaining inward stroke of the piston bringing the pressure and
temperature of residual steam up to admission values. In this way condensation
of
incoming steam by residual steam is avoided as well as in a steam-only uniflow
engine.
Efficiency, however, can be better than a uniflow steam engine because unlike
a standard
uniflow steam engine, where the piston surface adjacent the exhaust valve
cannot be
heated, here the entire adjacent cap 20 (Figs. 1, 2, and 18-22) above steam
chest 46 can
be steam jacketed in any engine application that would benefit from it and
thus heated
continuously by the steam chest 46 so as to prevent chilling the incoming
charge of
steam from producing a loss in power. The steam exhaust valve, piston,
cylinder and cap
20 are dimensioned, constructed and arranged to provide an effective zero
clearance in
chamber 44 and a gain in the mean Rankine cycle temperature as will be further
described below.
The mechanism for operating valves 200 and 202 will now be described with
reference to Figs. 18-20 and 22. These Figures show how steam admission poppet
valves
200 and 202 are placed in series relationship and are enclosed within piston
14 at the
upper and lower ends of inter-valve steam inlet passage 46a that extends
through the
cylinder cap 20 with valve 200 at its upper end and valve 202 at its lower end
so that
steam is admitted from the steam chest 46 into the steam expansion chamber 44
only
when both of the valves 200 and 202 are open. The stems of valves 200 and 202
extend
inwardly through the steam chest 46 within valve guides 204 and 206 (Fig. 20)
43
CA 02701481 2010-04-27
respectively, the former being threaded into the lower end of the passage 46a
by screw
threads 46b (Fig. 20). Both guides 204, 206 are not shown in Figs. 18 and 19
so that the
valves can be clearly seen. The guides 204 and 206 are sealed at their lower
ends within
openings provided in the crankcase 21 as shown at 208 and 210 (Fig. 20). The
stems of
valves 200 and 202 can be provided with shallow vertically spaced apart
circular grooves
201 to serve as a labyrinth seal. It should be noted that the valves 200 and
202 have
beveled valve face seals that are smaller in diameter at the outer end of each
valve,
providing distal seals, so that the valves open by being withdrawn inwardly,
away from
the steam expansion chamber 44 toward the crankshaft 30. Therefore, instead of
acting
as push rods, the stems act as traction rods to open the valves. By using
series-related
inwardly opening poppet valves 200 and 202, a relatively short cutoff is made
possible
with very little throttling while avoiding excessive G forces and cam stress
at high
RPM's. Consequently, the valve arrangement within the cap located inside of
the
combustion piston 14 as described makes it possible to accomplish a short
steam cutoff
in a high RPM combined cycle engine that has a piston speed compatible with
the piston
speed commonly experienced in an ordinary internal combustion car or truck
engine that
lacks thermal power recovery. At 3000 RPM each stroke lasts 0.01 second so
that a 10%
cutoff requires steam admission valves to open and close in 1/1000 second.
However, by
placing a pair of series-related inwardly retractable steam inlet valves in a
cylinder cap
that is slideably enclosed within the piston, it is possible to provide a
variable steam
cutoff even with short cutoff periods of 10% or less in a double-acting
combustion/steam
dual cycle engine in which steam is recompressed to admission pressure.
Operation of the valves 200 and 202 is accomplished by camshafts 212 and 214
through the action of centrally pivoted valve rockers 216 and 218 which open
the valves
by retracting them, i.e. moving them proximally (away from the steam expansion
44
CA 02701481 2010-04-27
chamber 44) so that during operation steam enters the steam expansion chamber
44 from
the steam chest 46 which is at a throttle pressure of say 1,000 p.s.i. Thus,
steam is
admitted when valves 200 and 202 are withdrawn responsive to a downward
movement
of the central ends of the rockers 216 and 218 against spring retainers 220
and 222
connected to the stems or traction rods of valves 200 and 202 which are
normally held in
the closed position by steam pressure and compression springs 224 and 226. The
distal
face seals of valves 200 and 202 can also permit the valves to open for
relieving excess
pressure whenever steam in chamber 44 is recompressed enough to exceed the
closing
force on valves 200 and 202 caused by steam pressure within steam chest 46
plus the
spring force. As described below and shown in Fig. 24, an auxiliary clearance
volume
chamber 352 similar to that in Fig. 2 is provided in the cylinder cap 20. The
pressure
balancing ability of valves 200, 202 in cooperation with the auxiliary
clearance volume
chamber 352 prevent excessive recompression counter-torque on the crankshaft
during
starting and idling. It will also be noticed that the steam inlet valves 200
and 202 are
easily moved by the camshafts 212, 214 due to the tendency for recompression
pressure
from above to open them, thus helping to balance the closing pressure from the
steam
chest 46. The steam admission valves 200 and 202 can therefore be opened with
a force
in the same range as that in a conventional internal combustion engine even
when steam
exerts a closing force of 500 lbs. or more on the valves.
The control of one or more steam inlet valves in accordance with the invention
is
used to continuously regulate the charge of steam admitted into the clearance
volume at the
beginning of each power stroke. While engine valve timing (the angle from the
bottom
center position of the crankshaft to the point at which the steam inlet valve
opens or closes)
and valve lift can be varied in any of several known ways in the present
invention, whatever
method is used should provide a regulated change in the mass of steam
admitted. Therefore,
the terms "valve control" or "valve cutoff' are used broadly herein to include
valve phase
CA 02701481 2010-04-27
shifting, valve lift changing, as well as changing the valve timing either
separately or
together for regulating the mass of steam that is admitted each cycle of
operation.
Valve control for regulation the mass of steam, admitted each stroke by means
of
valve phase angle control in place of a three-dimensional cam 64 (Fig. 2) will
now be
described with reference to Figs. 18-23. The left camshaft 212 runs at
crankshaft speed
through its connection by drive chain 230 and a cooperating sprocket 30a at
one end of
the crankshaft 30 and an equal size sprocket 232 at the end of camshaft 212.
The right
camshaft 214 is also driven via a sprocket 246 and a shaft 244 rotating at
crankshaft
speed, in this case through a variable valve timer, specifically, a phase
shift control 234
or other suitable valve timing device, variator or phaser of known
construction for
regulating the mass of steam admitted during each outward stroke of the
piston. While
two series related valves are preferred, the control 234 can be connected to a
single valve
in place of two valves 200 and 202 arranged to control the quantity of steam
admitted
each stroke. In Fig. 18 an electric servomotor 236 or other actuator is
connected for
shifting a positioning fork 238 parallel to camshaft 214 any selected distance
so as to
slide a helical gear 240 axially of a cooperating helical gear 242 at the end
of camshaft
214 which is engaged with the gear 240. It will be seen that the left end of
the
positioning fork 238 extends into a circumferential groove in a boss at one
end of gear
240 so that the helical gear 240 can rotate freely as the positioning fork 238
slides the
gear 240 during operation to any selected position from one end of helical
gear 242 to
the other. It will also be seen that the gear 240 is supported for sliding
movement on the
axially splined shaft 244 which is driven at crankshaft speed by the chain 230
via the
sprocket 246 so that gear 242 turns slightly relative to gear 240 as gear 240
moves
axially thereby operating valve 202 earlier or later in the cycle. The
electronic engine
management computer or controller 305 (Figs. 14, 18, and 25) is wired to
control the
46
CA 02701481 2010-04-27
operation of servomotor 236 to regulate the position of gear 240 and thereby
provide the
optimum steam charge, i.e. cutoff so as to achieve the lowest specific fuel
consumption
as previously described in connection with Figs. 1-17. The electronic motor
controller
305 itself is described in more detail hereinabove. The I.C. valves 36 and 38
can be
operated by an overhead camshaft (not shown) or in a lower cost engine, they
can be
operated from steam camshaft 212 running at half crankshaft speed by providing
each steam
cam on camshaft 212 with two similar steam lobes 180 apart. The timing of
valve 36 can
also be coordinated with valve 202 by an I.C. pushrod from camshaft 214 to a
valve rocker
(not shown) for valve 36.
Refer now to the graphs shown in Fig. 21 which illustrate valve lift vs. time
to
depict the opening and closing of valves 200 and 202. During operation, when
the
electronic engine management controller 305 signals the servomotor 236 to
shift the gear
240 a selected distance in a direction that spaces the opening of valves 200
and 202
further apart, there is an incremental rotation of shaft 214 due to the axial
movement of
gear 240 causing the phase of shaft 214 to be shifted, thereby altering the
overlap of the
valves so as to provide a short steam cutoff A as seen in the upper graph.
Delaying the
opening of the valve 202 as shown in the lower graph will provide a longer
steam cutoff
B by spacing the opening and closing of valves 200 and 202 closer together. In
this way,
by advancing or retarding the opening and closing of valve 202, a continuous
infinitely
variable steam cutoff is provided to control the charge of steam supplied to
each cylinder
for optimizing thermal power recovery and specific fuel consumption of the
combined
cycle engine 10 throughout operation.
During operation, the valves 200 and 202 are normally held shut by the
operating
steam pressure acting on the lower faces of the valves plus the spring force.
However, it
is desirable to have the recompression pressure in chamber 44 developed by the
47
CA 02701481 2010-04-27
downward movement of the piston be the same as the operating pressure even
when the
steam operating pressure is varied by changing the throttle setting. The
maximum
operating pressure, i.e. the generator or boiler pressure of the power plant
can be selected
first, for example, 1000 p.s.i. The clearance volume in the chamber 44 is then
set so that
the recompression pressure that results when the piston descends equals the
pre-selected
operating pressure of the steam generator, namely, 1000 p.s.i. However, when
the
throttle pressure is reduced below 1000 p.s.i., power is wasted during the
recompression
stroke. Consequently it is desirable to keep the recompression pressure in the
clearance
space equal to whatever the throttle pressure happens to be.
To accomplish this, in case excessive recompression occurs in the clearance
volume of chamber 44, a resulting drop in efficiency can be prevented by an
optional
pressure-regulating valve 360 (Fig. 24) which will now be described. Threaded
into
opposite ends of a passageway 361 that extends through the cylinder cap 20
from
chamber 44 to chamber 46 is an annular valve seat 362, and an annular spring
retainer
363 enclosing a relatively light spring 364 with enough force to reliably seat
a
frustoconical valve element 365 onto the seat 362. The clearance volume
between the
cap 20 and the piston 14 at bottom dead center is determined so as to
recompress residual
steam up to the maximum operating pressure, i.e. the boiler pressure, e.g.
1000 p.s.i.
However, throughout operation, whenever the throttle pressure is less than
1000 p.s.i.,
the valve 360 will open near the end of the recompression stroke equalizing
the pressure
by allowing excess steam to pass out of the clearance space thereby preventing
power
losses due to work performed in recompressing residual steam above the
throttle
pressure. The valves 200 and 202 will then be opened only by the valve
retractors and
not by a pressure differential across them.
48
CA 02701481 2010-04-27
During operation, the effective clearance volume within the steam expansion
chamber 44 can also be varied as explained above by the auxiliary clearance
volume
chamber 45 (Fig. 2) or 352 (Fig. 24) and if valve 360 is used, the maximum
recompression pressure that can be developed will also be controlled by the
regulating
valve 360 so that when the piston reaches the end of its inward stroke, the
recompression
pressure will be equal or almost equal to the throttle pressure in chamber 46
whatever its
value. Thus, at the opening of the admission valves 200 and 202 at or near
BDC, no
steam flows into chamber 44 of the cylinder because the cylinder is already
filled.
Consequently, no steam mass is consumed just to fill the clearance volume. The
result is
an effective zero clearance. When steam does flow into the cylinder, its mass
is totally
consumed by admission and expansion work. Steam is therefore more efficiently
utilized, thus improving efficiency of the engine. Also, as recompression
occurs, the
temperature of the recompressed steam will rise up to or above the admission
supply
temperature. The recompressed steam mixes with the supply steam admitted
through
valves 200, 202 resulting in a steam temperature at cutoff that is most
preferably greater
than the supply temperature thereby producing a gain in the mean cycle
temperature and
when the mean cycle temperature is elevated, the efficiency of the engine is
enhanced.
These two events of course occur at the expense of the work of recompression.
However,
thermodynamic analysis indicates that there is a net improvement in efficiency
due to an
effective zero clearance and an increase in the mean cycle temperature which
produces
an increase in output that is greater than the fraction of the recompression
work that
cannot be recovered during the expansion stroke. It can therefore be seen that
the present
invention is able to provide a dual cycle internal combustion steam engine
having an
effective zero clearance in the steam expansion chamber 44 as well as the
capacity for
achieving a mean cycle temperature gain thereby assuring a higher level of
Rankine
49
CA 02701481 2010-04-27
efficiency. It should be noted that if valve 360 is used, it will open
whenever the
pressures in chamber 44 is slightly above that in chamber 46 while in the case
of valves
200, 202, the opening force will be significantly greater than the net closing
force whenever
the throttle is partly closed.
The retractable series valve arrangement of the valves 200 and 202 in the cap
20
which is in turn enclosed within of the internal combustion piston 14 brings
steam valve
acceleration and operating rates to reasonable levels even if a short cutoff
period is
required as optimum efficiency is achieved while at the same time the closing
force on
the valves is balanced in large part through the recompression pressure
developed as the
piston descends acting in cooperation with the auxiliary clearance chamber 352
(Fig. 24).
The series-related valves 201 and 202 are, moreover, effective in overcoming
the valve
acceleration and stress problems within the cutoff range (5% to 12%) that is
of interest
because it will guarantee a steam use rate which is commensurate with the
water
evaporation rate of the steam generator.
EXAMPLE
CALCULATION OF POWER RECOVERED
A one-cylinder 564cc (34.4 cubic inches), 4-cycle internal combustion engine
(bore 96mm x stroke 78mm) producing 13.5 horsepower (572.6 Btu/min) at 3000
rpm
and exhibiting a brake efficiency of 21 % will require an energy input rate of
2526.7
Btu/min and will reject 2154.1 Btu/min to the steam generator 104. For a
boiler
efficiency of 70%, which is merely average for modern monotube steam
generators,
1507.8 Btu/min will be transferred to heat the feed water. The sustained
evaporation is
the ratio of the waste heat recovered to the enthalpy change L h as the feed
water (or
steam) passes through the steam generator. At 800 psia and 800 F with
saturated feed
CA 02701481 2010-04-27
water at 14.7 psia and 212 F, the enthalpy change is 1218.5 Btu/lb (1398.6-
180.07). The
resulting sustained evaporation rate for supply steam mS is then 1.24 lb/min.
:
waste heat recovered 1507.8 Btu/min _ 1.24 lb/min
Ms h 1218.5 Btu/lb
Rankine cycle analysis based on the internal energy changes of non-flow
process within
a 24 cubic inch (399cc) displacement high compression uniflow steam expansion
chamber 44 (e.g. Figs. 18-25) with a 7% cutoff will yield a brake efficiency
of 18.9 %
with supply conditions at 800 psia and 800 F. This level of performance is
consistent
with an engine friction torque of 0.10 foot pounds per cubic inch displacement
and pump
work necessary to raise the feed water pressure from 14.7 psia to an operating
pressure of
800 psia. Steam consumption under these constraints is 1.11 lb/min. This is
consistent
with the evaporation rate of 1.24 lb/min. Power delivered by steam = .189x
1507.8
Btu/min/42.41 Btu/min/hp = 6.72 horsepower. The combined performance of both
internal combustion and steam engine therefore represents a 49% increase over
the 13.5
horsepower internal combustion engine acting alone. Regenerator use and heat
recovery
from unburned fuel provides a still greater increase. At a 75% boiler
efficiency consistent
with monotube steam generators tested in cars, the brake Rankine efficiency
would be
20.4% producing 7.25 HP from steam which is a 54% increase in power under the
same
conditions. When used in a hybrid electric installation, even greater
economies are
obtained.
CONSOLIDATED VALVES
Refer now to Fig. 22 which illustrates a modified form of the invention having
valves that operate in a manner similar to valves 200 and 202 except that in
this example
a pair of admission poppet valves 200a and 202a are concentrically arranged so
that the
51
CA 02701481 2010-04-27
valve 200a slides within a central bore in a tubular poppet valve 202a. The
lower valve
202a is mounted to slide within a valve guide 260 which is secured at 262 to
the
crankcase 21 to provide a steam-proof seal. Valves 200a and 202a are normally
held in
the closed position within the cap 20 inside of the combustion piston by
compression
springs 207 and 209. In this case, the valves 200a and 202a are operated by
vertically
spaced rockers 270 and 272 respectively, which engage vertically spaced apart
cams of
camshafts 212a and 214a, the timing of which is controlled in the same manner
as
described in connection with Figs. 18-21. Thus, the phase of the camshaft 214a
is
advanced or retarded with respect to camshaft 212a to thereby regulate the
cutoff of
steam through the sequential operation of valves 200a and 202a which together
permit
the admission of steam through an intervalve passage 203 into the steam
expansion
chamber 44 when both admission valves 200a and 202a are open. This enables the
steam
cutoff to be varied throughout operation as determined by the electronic
engine
management controller 305 through variable cam positioning, namely by changing
the
phase angle of shaft 214a relative to shaft 212a as described in connection
with Figs. 18-
21. The concentric valves provide the advantage of minimizing the size of
passage 203
within casing 205 thereby assuring better control of the volume of steam
admitted,
especially at a short cutoff as well as a straight passage 302 and enlarged
steam chest
volume since only one valve guide is needed.
ENGINE INSTALLATIONS AND STEAM PRODUCTION
The engine 10 can be installed and used in various ways. For example, the
engine
can be connected directly to the wheels of a vehicle as described above (see
Fig. 12).
When mechanically connected to the wheels to power the vehicle, the engine
will operate
at varying speeds and loads that result from continuously changing driving
conditions. In
52
CA 02701481 2010-04-27
such an installation, the invention provides operational flexibility in spite
of rapidly
changing driving conditions as well as an ability to function at I.C. engine
speeds, and a
way of matching the engine displacement with the changes in steam generator
output as
described hereinabove and in the prior related pending application Serial
Number
12/075,042 (Publication No. US 2008/0216480 Al), published September 11, 2008
which is incorporated herein by reference, now patent . However,
instead of powering the wheels, the engine can also be used in a hybrid
vehicle as a
constant RPM, reduced size, battery charging module to increase the range of a
vehicle.
In this application, the engine 10 is operated at an optimum fixed RPM and
load while
connected to an electric generator and is run for extended periods of time to
recharge a
battery and/or an ultracapacitor, i.e. as a recharging module and/or to power
an electric
drive motor that is coupled to the wheels of the vehicle. This enables the
engine to
operate as efficiently as possible at a higher speed for extended periods with
the
dimensions of steam engine components and their operating parameters chosen to
provide
the lowest specific fuel consumption for the particular internal combustion
assembly
components used in any given engine.
A power plant installation especially useful as a battery charging module to
extend the driving range of a hybrid I.C./electric vehicle but which could
alternatively be
connected mechanically to power the drive wheels is shown in Fig. 25 wherein
the same
numerals designate corresponding parts described above and in application
12/075,042
filed March 7, 2008. For simplicity and clarity of illustration, the complete
engine
coolant circuit within the engine and the steam generator 100 of application
12/075,042
has not been shown in Fig. 25 but both can, if desired, be constructed as
previously
described. In Fig. 25 it will be seen that the engine 10 is connected
mechanically by shaft
548 to an electric generator 550 which is wired at 552 to a power supply 554
that
53
CA 02701481 2010-04-27
provides electric current to storage batteries 557 and/or ultracapacitors 559
through
conductor 556 under the control of the electronic central engine management
computer
305. Current from the power supply 554 can also be provided through conductor
558 to
an electric motor generator 560 which is connected by shaft 562 to the drive
wheels 561
of a vehicle such as an automobile, truck, locomotive, or propeller of an
aircraft. Thus,
during operation, the engine 10 is run at an optimum speed and load which is
typically at
a fixed RPM selected for recharging the ultracapacitor 559 and battery 557
when
required and/or to provide electric power to the motor 560 which can be
supplemented
by power from the ultracapacitor 559 and/or battery 557 whenever additional
power is
needed. When the battery is charged above a set level, the engine 10 can be
turned off by
the motor controller 305 and the electric motor 560 then operated by the
battery and/or
ultracapacitor either separately or together. In such an installation, the
vehicle is run initially
on current from the battery 557 and/or ultracapacitor 559 while the engine 10
is used
primarily as a back-up battery recharging device to increase the range of the
vehicle.
During braking, electric motor 560 acts as a generator for recovering momentum
by charging the battery and/or capacitor under the control of the CEM computer
305 for
later moving the vehicle ahead.
In place of an ordinary liquid coolant system used in a conventional I.C. car
engine, heat can optionally be removed from the combustion chambers of the
engine 10
by boiling the coolant, if desired, i.e. by evaporative cooling to produce
steam in the
cooling jacket 12b and 39. This can be accomplished, for example, by spray
cooling each
cylinder to generate steam. Evaporative spray cooling can be carried out as
disclosed by
D.A. Arais, et. al., SAE Technical Paper, 2006-01-1605 in which the cylinder
is cooled
by a spray of liquid coolant from several small parallel vertical pipes (not
shown) within
the jacket 12b surrounding the cylinder to produce steam. Compared to a
conventional
54
CA 02701481 2010-04-27
water-cooled engine, evaporative cooling absorbs 970 additional Btus of heat
per pound
of water evaporated in the production of steam; the heat of vaporization. The
resulting
steam is able to accomplish much more efficient heating when the coolant is
fed to a
steam generator as steam rather than as hot water. This is explained by the
better match
between the temperatures of the energy sources; steam from the I.C. engine
block (in
place of hot water) and the engine exhaust gas used to power the steam
generator.
Evaporative cooling is carried out within the engine cooling jacket in a
manner that
maintains nucleate boiling so as to avoid exceeding a critical heat flux which
can result
in the formation of an insulating vapor blanket layer within the liquid at the
hot metal
surface allowing the temperature of the metal to jump suddenly as runaway
heating takes
place with potentially harmful results such as engine knock (detonation).
Spray cooling
is one way of avoiding hot spots. Another way is to impart vibratory movement
to the
liquid by attaching an ultrasonic vibrator to the cooling jacket. Kwon, et.
al., were able to
substantially increase the critical heat flux using a 40 kHz ultrasonic
vibrator (see
Experimental Study On CHF Enhancement In Pool Boiling Using Ultrasonic Field,
J.
Ind. Eng. Chem., Vol. 11, No. 5 (2005 pg. 631-637). Evaporative cooling can
also be
carried out as described in U.S. patents 3,731,660; 4,565,162; and 7,421,783
or as
described in Ap, et al., New Components Development for New Engine Cooling
System
VTMS4, 1999, Paper C 543/047/99. See also SAE Paper by Chamfreau, et. al., No.
2001-01-1742.
Refer again to Fig. 25 which illustrates an example to show how the invention
can be installed and used as a constant speed, high RPM battery recharging
module. In
this example of the invention, engine cooling and final steam production are
integrated in
series by circulating a single fluid in a closed loop to serve as an engine
coolant as well
as a working fluid in the engine. Thus, the fluid is heated first in the
combustion chamber
CA 02701481 2010-04-27
cooling jacket (12b and 39 Fig. 1) preferably to form steam by evaporative
cooling as
described above. The steam then flows to the generator/superheater 104b-104
where it is
heated further by combustion exhaust gas to provide superheated steam under
high
pressure that is supplied through the throttle T to the steam expansion
chambers 44 of the
engine 10 below the pistons 14. By running the cooling chambers 39, 12b (Fig.
1) at a
high enough temperature to evaporate the coolant within the cooling jacket
itself, steam
collects at a controlled pressure above atmospheric pressure in the chamber
500 just
above the combustion chambers 34. In operation, the steam flows out through a
steam
duct 504 to a pressure regulator valve 506 which maintains a predetermined
pressure
within the engine 10. For example, at 25 psia, saturated steam produced in the
engine
will be at a temperature of 240 F. Once the steam has reached the
predetermined
pressure established-by valve 506, it will then pass through supply line 508
to
countercurrent flow heat exchanger or regenerator 106 where low-pressure steam
exhausted from the steam expansion chambers of the engine 10 through line 52
to line
114 enters the heat exchanger 106, flowing in the opposite direction thereby
transferring
a part of its heat load to the low temperature steam formed in the engine
cooling jacket
39, 12b (Fig. 1). Pressure in the steam generator and superheater 104 is
maintained by a
feed pump 511 in line 510. From the heat exchanger 106, the steam which has
now been
heated to a temperature approaching the temperature of exhausted steam, flows
through
pump 511 into the superheater 104 which has been extended by a pre-heater
section 104b
to a total length of about 6 feet or more and contains additional heater coils
130 that in
the figures are depicted as a single spiral but which can consist of a total
of 58 or more
pancake coils 512, e.g. of 5/8" steel tubing connected end to end and spaced
about 1'/4
inches on centers. Each pancake coil 512 can be about 60 inches long to
provide a total
of about 290 feet of tubing (52 sq. ft. of heating surface) providing a 24 HP
steam
56
CA 02701481 2010-04-27
generator in which little power is lost due to backpressure. Superheated steam
that is
formed in the superheater 104 flows as described above through the throttle T,
then
through the high-pressure steam supply line 49 and valves V to the steam chest
46 or
directly to valve 202 (see Fig. 23) then to the steam expansion chambers 44 to
power the
engine as described previously. The low-pressure exhaust steam from the heat
exchanger
106 after having transferred its heat load to the steam from the engine
cooling jacket is
pumped from line 514 by a compressor 516 through line 518 to a condenser 520
which is
maintained by the compressor 516 at an elevated pressure substantially above
atmospheric pressure so as to achieve a high rate of cooling by the condenser
520 owing
to a substantial temperature difference between the ambient air passing
through the
condenser and the pressurized steam entering the condenser. Condensed steam
collects at
the bottom of the condenser 520 where it drains into a storage tank 164. The
pressurized
condensate in the storage tank 164 flows through a line 522 to a pressure
regulator valve
524 which maintains the high pressure in the condenser 520 and in storage tank
164.
From valve 524, condensate flows at a relatively low pressure through a
feedwater line
526 to a countercurrent flow heat exchanger 528 where it can be preheated
under certain
operating conditions by diverting the flow from line 518 by valves 530 and 532
through
the heat exchanger 528 when steam in line 518 is at a significantly higher
temperature
than the feedwater entering through line 526. From the heat exchanger 528, the
feedwater is pumped by a feedwater pump 534 through line 536 back to the
engine
cooling jacket 12b to complete a closed circuit where it is again evaporated
to form
steam within the cooling jacket 12b and 39 of the engine 10.
During operation, the I.C. exhaust gases passing through exhaust pipes 141-144
into the exhaust manifold 104-104b which functions as a superheater, pass out
through
exhaust pipe 103. As noted above, spent steam exhaust from engine 10 is
carried by pipe
57
CA 02701481 2010-04-27
52 through line 114 to the regenerator 106, then to condenser 520 and through
line 536 to
the cooling jacket of engine 10. Coolant leaving the engine as steam through
duct 504
picks up the residual heat from the exhausted steam in regenerator 106 before
entering
the generator/superheater 104-104b. Superheated steam from the superheater
section 104
passes through the throttle T to the high-pressure steam line 49 into the
engine 10 while
the steam cutoff is regulated by engine controller 305 to maximize efficiency
as
described above and in application Serial Number 12/075,042, now patent
. In Figs. 18-25 the steam charge cutoff is controlled as previously
described by changing the phase angle of camshaft 214 to optimize the specific
power
output recovered from steam. One or more engine operating variables including
the fuel
consumption rate can be used as inputs to electronic computer 305 for
controlling the
operation of the throttle T and the cutoff servo 236 (Fig. 18) as previously
described.
Rankine cycle operation is most effectively optimized by regulating both steam
cutoff
and throttle control.
As noted above, operating conditions experienced by the engine vary greatly
depending upon how the engine is used and applied to the load, whether
connected to
drive the wheels of a vehicle or run at a constant RPM for extended periods as
in
powering an electric generator that is used to charge a battery. In the second
application,
and particularly when accompanied by the use of an afterburner for consuming
unburned
exhaust constituents as shown in Figs. 9A, 9B and 25, the engine 10 may
produce
relatively high temperature exhaust gas, e.g. at 700 C to 900 C for long
periods of time.
When run continuously in this manner, the temperature of the steam being
supplied to
the engine may reach a level sufficiently high that the cylinder cap 20 will
become
overheated, that is to say, if a high degree of superheat is provided over an
extended
period, it is possible to heat the upper surface of the cap 20 enough so that
heat carried
58
CA 02701481 2010-04-27
away in the steam exhausted from the cylinder outweighs the saving provided by
heating
the cap 20 to prevent the cap from chilling and thus reducing the enthalpy of
each fresh
charge of high-pressure steam. In such a case, where there is excessive heat
loss from the
cap, the cap can be provided with either no external heating or a controlled
degree of
external heating. One way this can be accomplished is through the use of a
heat
insulating layer inside the cap such as an optional porous mineral liner 17
(Figs. 19 and
20) having relatively low thermal conductivity to insulate and thereby limit
the heat flux
from steam chest 46 to the upper surface of the cap 20 to a degree that
optimizes heat
transfer to chamber 44. Thus, according to this feature of the invention, the
rate of heat
flow from the steam chest 46 through the cap 20 is limited, i.e. controlled.
It can be seen
that the optional insulating layer 17 of Figs. 19 and 20 reduces heat transfer
through the
cylinder cap 20 while the valve compartment 201 containing valves 200 and 202
is not
covered by the optional insulating layer 17 and therefore remains at the
temperature of
the steam chest 46. The controlled heating rate can be determined by selecting
the
thickness and thermal properties of the insulating layer 17 to optimize the
transfer of heat
to the top of the cap 20 as judged by the power output or specific fuel
consumption of the
engine. Accordingly, it is an optional feature of the invention to provide a
cylinder cap
inside an internal combustion piston with either no external heating or
controlled
external heating of the cylinder cap 20. Thus, heat control is accomplished in
this case by
optionally providing a heat source, e.g. the steam at 46 in heat-conductive
relationship
with the cylinder cap through the heat transfer barrier, e.g. layer 17,
between the cap and
the heat source for reducing heat flow through the cylinder cap to the steam
expansion
chamber 44 to a level below that produced by steam jacketing the cylinder cap
as in Figs.
1-8. On the other hand, in applications when the operation of the engine 10 is
typically
intermittent as in a hybrid gas engine/electric vehicle such as the Toyota
PriusTM car
59
CA 02701481 2010-04-27
where the engine is mechanically connected to the drive wheels, heat flow to
the cap 20
can be uncontrolled so as to maximize heating of the cap by high-pressure
steam in the
steam chest 46. However, if the engine 10 is connected instead only to drive
an electric
generator at a constant relatively high speed for recharging batteries, etc.,
as in the G.M.
VoItTM car or any other application in which overheating of the cap could
reduce steam
efficiency, heating from a source of heat such as the steam chest 46 can be
completely
removed or reduced in any suitable way by controlling heat flow to the top of
the cap 20 to
an optimal level based on power output.
Refer now to Fig. 23 which shows the engine 10 constructed without the steam
chest 46. Instead, steam from the superheater 104 flows through a feed pipe
49a directly
to valve 202. In accordance with the optional feature of the invention shown
in Fig. 23,
the cap is either not heated externally at all or, if desired, limited heat is
optionally
transferred in a different way to the cap 20 at a controlled rate. In this
embodiment,
steam flows from supply line 49 through the vertical pipe 49a directly to the
valve 202,
i.e. without being allowed to enter the chamber inside cylinder 12 below cap
20. The
space 46 inside the cylinder 12 below the cap 20 can be sealed at the lower
end of the
cylinder 12 as shown in Figs. 19 and 20 or open at 51 and 53 (Fig. 23) to the
interior of
the crankcase at 31, if desired with the crosshead guide 20a held securely in
place, e.g.
by braces as at 55 and by pipe 49. Consequently, there is no significant
direct heating of the
cap; only incidental heating in the area of the steam inlet valves. However,
to accomplish
controlled heating when some heat is desired, limited heat can optionally be
supplied to
the top of cap 20 through a small steam pressurized supply passage 57 having
an inlet 67
connected to steam supply pipe 49b and fed via distribution channel 59 within
the cap to
a few small dead ended radial chambers 61 connected to the cap or within the
cap 20.
The chambers 61 provide heat at a controlled rate to the top of the cap; the
temperature
CA 02701481 2010-04-27
gradient and maximum flux in Btu's/hr/ft2 being significantly less than when
the entire
lower surface of the cap is steam jacketed by the steam chest 46 as shown in
Figs. 1-4.
The size and number of the chambers 61 is chosen to provide a heat flux that
maximizes
the specific power output of the engine 10. The pressure-regulating valve 360
(not shown in
Fig. 23), if used, is placed in this case between chamber 44 and pipe 49a.
Refer again to Fig. 24 which illustrates an alternate optional auxiliary
clearance
volume chamber at 352. Except for the removal of condensed steam, the
auxiliary
clearance volume 352 of Fig. 24 functions in the manner of the auxiliary
clearance
volume chamber 45 of Fig. 2 as described above and in pending application S.N.
12/075,042, now patent by admitting steam through a throttling duct
350 of selected length and diameter. The auxiliary clearance volume chamber
352 is
located in the cylinder cap 20 so that the throttling duct 350 communicates
with the
steam expansion chamber 44 at the top of the auxiliary clearance volume. The
auxiliary
clearance volume 352 like 45 (Fig. 2) reduces the pressure in steam chamber 44
at lower
engine speeds as previously described. During operation when the throttle
pressure is
lowered, engine speed is, of course, reduced. Hence the auxiliary clearance
chamber 45
or 352 can be used to automatically lower cylinder pressure to the reduced
throttle
pressure thereby eliminating the need for valve 360. In Fig. 2, gravity will
drain
condensate (if any) from chamber 45. Because of the high operating
temperatures, no
condensate should accumulate in chamber 352. However, if condensate does tend
to
form, a suitable thermostatic drain valve 356 can be used to remove the
condensate
through a drain 354 by opening when the engine is cold and closing when the
engine is
hot.
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CA 02701481 2010-04-27
CHARACTERISTICS AND UNEXPECTED RESULTS
From the description of Figs. 1 and 2 it can be seen that the invention is
capable
of providing even better protection against the chilling of the metal surfaces
by low-
pressure exhaust steam than a conventional steam-only uniflow engine. A
uniflow steam
engine exhausts steam adjacent to the piston head and there is no way to heat
the surface
of the uniflow piston head positioned adjacent to the ports through which
steam is
released. By contrast, in the present invention low-pressure steam is
exhausted at the cap
20 comprising the steam cylinder head, exactly where heat can be supplied to
the steam
expansion chamber by the adjacent steam chest 46. Consequently, for certain
applications, especially if start and stop operation and idling is essential,
the present
invention is able to provide better heat loss protection and therefore greater
Rankine
efficiency than that in a standard uniflow steam engine. The invention can
provide better
protection against power losses during operation in part through enhanced
control of the
steam mass admitted each stroke, e.g. by the provision of one or a pair of
steam
admission valves in the cap that open into the steam chest and in part through
special
exhaust valving. which in one preferred form comprises a double aperture
automatic
exhaust valve having ports in both the piston skirt as well as in the cylinder
wall which
are covered when steam is admitted but which open adjacent the cylinder cap
when
aligned with one another while the steam chamber is in an expanded state.
Additionally,
the invention provides a combination I.C./steam engine in which residual steam
within
the piston can be recompressed so as to provide as effective zero clearance
with a gain in
mean Rankine cycle. temperature. Moreover, low-pressure exhaust steam need not
be
ducted below the cylinder cap where it could cool the cap or the steam chest
which is
therefore able to jacket the entire lower surface of the steam cylinder cap in
any
installation in which it is advantageous to do so. In addition, the invention
provides an oil
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ring on the piston for stripping combustion products and excess oil from the
cylinder and
depositing it in an annular collection channel that is built into the wall of
the cylinder at a
point located in alignment with the oil ring when at bottom dead center and a
passage for
carrying excess oil and blow-by combustion products to the sump. Optionally,
the
collection channel holds one or more wiper rings that press centrally against
the outer
wall of the piston for capturing excess oil and blow-by combustion products
and ducting
them to the engine sump before they can reach the steam exhaust manifold.
Besides sealing the steam expansion chamber 44, the stationary cylinder cap
also
provides support for one or a pair of steam inlet valves, reduces the steam
clearance
volume between its upper surface and the confronting inner wall of the piston,
defines
the top of the steam chest, can heat the steam exhaust area and supports
packing around
the piston rod to prevent the escape of steam. By providing a piston with a
skirt that is in
sealing and sliding relationship between both the steam cylinder head and the
cylinder
walls, the added length of the cylinder needed to accommodate steam is
minimized since
the steam chamber and I.C. chamber occupy the same space. In addition, the
invention
provides an improved system and sequence of heat transfer devices for
efficiently
recovering waste combustion heat from the I.C. engine assembly.
In summary, the invention provides the following benefits and unexpected
results
among others: 1) efficiency in a combined cycle engine as good or better than
a uniflow
steam engine while adding as little as possible length to the engine
cylinders, 2) the
steam expansion chamber occupies the same space as the combustion chamber of
the I.C.
engine, 3) the steam expansion chamber is heated by direct thermal transfer
from the I.C.
engine combustion chamber, 4) the burnt oil from the I.C. chamber as well as
blow-by
combustion gasses are kept separate from the steam expansion chamber and the
exhausted steam, 5) the exhaust valve need not exhaust low-pressure steam
below the
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steam cylinder head where it could cool incoming steam or interfere with steam
jacketing
of the steam cylinder head, 6) clearance volumes can be as small as desired
since the
opposing surfaces of the piston and cylinder cap are both flat, 7) the exhaust
valve when
opening sweeps adjacent steam and moisture if any out of the cylinder, then
closes so
that the cylinder pressure is brought up to or very close to the steam
admission pressure
thereby preventing steam at atmospheric pressure from washing heat from the
cylinder
walls and head, which if it occurred could reduce power due to the
condensation of
incoming steam, 8) piston walls are exposed to exhaust steam pressure only for
the short
period that the exhaust port is uncovered, whereupon the pressure and
temperature
immediately begin to rise so that at the end of the down stroke the
temperature and
pressure is such that an effective zero clearance is achieved with a gain in
mean Rankine
cycle temperature whereby efficiency is improved and the incoming steam meets
relatively hot surfaces thereby preventing cooling of the incoming charge, 9)
unlike a
standard uniflow steam engine, the metal surface at the end of the steam
chamber
adjacent the exiting steam can be externally heated when it is advantageous to
prevent
exhaust steam at atmospheric pressure from chilling or condensing the incoming
charge of
steam, 10) a short steam admission cutoff can be provided without excessive
valve and cam
stress, 11) evaporative cooling improves steam production efficiency, 12)
controlled heating
or no heating of the cylinder cap reduces thermal loss in the exhausted steam,
13) steam
recompression pressure can be limited to throttle pressure, 14) the engine can
be
mechanically connected to the wheels of a vehicle or used to charge a battery,
and 15)
momentum energy of a vehicle can be stored and used later.
Many variations of the present invention within the scope of the appended
claims
will be apparent to those skilled in the art once the principles described
herein are
understood.
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