Note: Descriptions are shown in the official language in which they were submitted.
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HIGH EFFICIENCY MULTICYCLE INTERNAL COMBUSTION ENGINE
WITH WASTE HEAT RECOVERY
This application claims priority based on United States Patent Application
12/492,773 entitled "HIGH EFFICIENCY MULTICYCLE INTERNAL COMBUSTION
ENGINE WITH WASTE HEAT RECOVERY" filed June 26, 2009 and United States
Patent Application 12/539,987 entitled "HIGH EFFICIENCY MULTICYCLE
INTERNAL COMBUSTION ENGINE WITH WASTE HEAT RECOVERY" filed
August 12, 2009, which are 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 piston 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 of
fossil fuel
decline and greenhouse gas emissions continue to rise. 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 No. 6,834,503) requires, in
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addition to the internal combustion 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 No. 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 Nos. 25,356 of
1910 and
28,472 of 1912 and U.S. Patent No. 1,324,183) and Mason U.S. Patent No.
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. Patent Nos. 2,402,699 and 2,943,608 reported tests showing a
thermal
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efficiency of 38.2% at 3450 RPM. A double acting hollow piston uniflow engine
is
described in Marks 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 engine for a car 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 as or better than in what is 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.
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
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that a 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 I.C. engine coolant and from engine exhaust gases. 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.
A further more specific object is to provide a multicycle engine in which
steam acts on
an internal combustion piston without the requirement for a piston rod, rod
seal or
crosshead while at the same time reducing the length of the cylinder.
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 to be considered 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 showing the piston at the top dead center
position.
Fig. 1 A is a partial sectional view taken through an end of the wrist pin 26
of Fig.
5 1 to show an alternate pin mount.
Fig. 2 is an exploded perspective view of the steam cylinder head or cap and
piston.
Fig 2A is a partial vertical sectional view of the piston of Fig. 2 taken
through the
center of one of the tongues 14t to show an optional modified form of piston
wall.
Fig 3 is a horizontal, cross-sectional view taken online 3-3 of Fig. 1.
Fig. 4 is a partial vertical sectional view taken online 4-4 of Fig. 1 with
the lower
cylinder head assembly or cylinder cap shown partly in side elevation.
Fig. 4A is a partial vertical crossectional view of the cylinder cap on a
larger
scale to show blow-by steam collection.
Fig. 5 is a vertical sectional view of a modified form of lower cylinder head
assembly or cap to show how steam can be supplied directly to the steam
admission
valve through a supply pipe.
Fig. 6 is a vertical sectional view similar to Fig. 5 showing optional
concentric
steam admission valves according to the invention and
Fig. 7 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 a
combustion
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chamber 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 valve 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 cutoff control, a camshaft is coupled for changing the
phase of a
single valve or a pair of steam admission valves in which the overlap is
varied, thus
providing continuous regulation of the steam cutoff to further reduce specific
fuel
consumption. These objectives are accomplished while combustion and steam act
on
opposite sides of each piston, yet, without the need for a piston rod, rod
seal, crosshead
or guide through the provision of a coupling between the piston and connecting
rod that
is positioned inward of the steam cylinder head or cap so as to reciprocate
within an
opening extending along the axis of the cylinder between the outer end of the
cylinder
cap and the crankshaft.
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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, is able to
provide an intermediate controlled degree of heating to minimize potential
power losses.
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 so
as 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. 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. 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, valving and heated
steam
exhaust area. There is also a provision for steam recompression to admission
pressure
inside of a combustion piston so as to achieve an effective zero clearance
volume and a
gain in mean Rankine cycle temperature along with a steam supply arrangement
that is
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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 without
the use of a
piston rod, crosshead or crosshead guide thereby reducing the reciprocating
mass.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
Refer now to the drawings in which the same numbers refer to corresponding
parts in several views. Shown diagrammatically in Figures 1-7 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. The piston 14 is provided with two compression
rings
and at the bottom an oil ring all marked 54. Near the lower end of the piston
are
circumferentially arranged exhaust openings 14b in the piston skirt.
Positioned above
and below the exhaust openings 14b are additional compression rings 54 (Figs.
1 and 2).
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. Any
suitable working
fluid such as water or water mixed with another fluid such as ethyleneglycol
or other
known working fluid can be used.
The steam cylinder head (Figs. 1-7) which is located within the piston 14
comprises a flat hub, disk or circular cap 20 having a top wall that may be,
say, 1/4 to 1/2
inch in thickness supported at the free upper end of a pair of integral
axially extending
laterally spaced arcuate (preferably semicircular) left and right legs 201 and
20r
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connected to a circular flange that is secured to the crankcase 21 by bolts 21
a. The disk
or cap 20 acts as a lower or inner steam cylinder head or end cap comprising
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 at an intermediate location that is spaced from its ends. 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 (Fig. 2A) to provide a hard piston ring contact surface.
Between the legs 201 and 20r on each side are axially extending opposed slots
20s
and 20t (Fig. 4) that provide an opening which traverses the cylinder 12
between the
cylinder cap 20 and crankshaft 30. In the present invention there is no piston
rod, rod
seal, crosshead or crosshead guide. Instead, to reduce the number of parts as
well as the
reciprocating mass and the length of the cylinders, there is provided on each
side of the
lower end of the piston 14 pair of opposed inwardly extending (toward the
crankshaft)
tongues 14t bored to hold the ends of a transversely extending wrist pin 26
(Figs. 1 and
3). The ends of the wrist pin 26 can be held in the bored openings 14e in each
tongue by
a pin for example extending through an opening 14f (Fig. 2) or a boss 24 can
be press
fitted through each tongue during assembly to hold the wrist pin 26 (Fig. IA).
Pivotally
supported on the wrist pin 26 is the upper end of a connecting rod 28 (Figs.
1, 3 and 4).
Thus the openings 14e serve to support the ends of the wrist pin which at
t.d.c. is located
just below the top of the cylinder cap 20 for coupling the connecting rod 28
to the piston
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14. It is preferred to have the connecting rod 28 offset toward one end of the
wrist pin 26
as in a Lancia engine and other engines to make possible the placement of a
steam
admission valve 48 and valve guide 31 closer to the center of the piston
(Figs. 3 & 4).
Due to the presence of the inner cylinder cap 20, the piston cannot be secured
to
5 the connecting rod before insertion into the cylinder. Instead, the piston
is placed on the
cap 20 in the b.d.c. position, the wrist pin then inserted through the
openings 14e as well
as through the bearing at the upper end of the connecting rod 28. The cap 20
and piston
can then be elevated into cylinder 12 and secured by the bolts 21 a to the
crankcase 21
before attaching the big end of the connecting rod 28 to the crankshaft 30 as
the wrist pin
10 is slid upwardly through the slots 20s and 20t between the legs 201 and
20r. If desired,
the flange at the lower end of the legs can be omitted and the lower end of
the legs
provided with screw threads so that the bottom end of the cylinder cap 20 can
be
threaded into the cylinder 12 instead of being bolted to it.
An alternate form of piston (not shown) has a separate lower cylindrical
threaded
segment just below the openings 14b that is screw threaded onto the skirt
during
assembly. The lower segment has the tongues 14t with aligned centrally
extending
integral bosses for the wrist pin similar to the bosses of an ordinary piston.
The steam admission valve 48 is an inwardly retractable poppet valve that
seals
the steam expansion chamber 44 by contact with a conical valve seat 48a (Fig.
5)
adjacent a port in an upper wall 22a of the cylinder cap 20 (Figs. 1 and 2).
The cap 20
also has a lower wall 22b defining a chamber 46 which serves as a steam chest
that is
supplied with steam through a passage 22d as a part of the leg 201. Steam is
supplied at
throttle pressure to the passage 22d through a pipe 49 that is threaded into
an opening
49a in the bottom of the leg 20r. The valve 48 is slidably mounted in a
central bore of a
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valve guide 31 which has an enlarged head threaded at 31 a into a lower wall
22b of the
steam chest 46. Attached to the lower end of the valve stem of valve 48 is a
stop 48b in
contact with one end of a valve rocker 64a that is operated in timed
relationship with the
rotation of a crankshaft 30 and cam 64 on a cam shaft 61 that can be advanced
or
retarded for controlling the steam admission and cutoff as described in
copending
applications S.N. 12/075,042 and S.N. 12/387,113 which are incorporated herein
by
reference. When advantageous in a particular design, the lower end of the legs
201 and
20r can each be provided with a slot or pocket 20m (Figs. 2 and 4) to act as
recesses if
needed to prevent the connecting rod from contact with the side wall.
It will be seen in Fig. 2 that the steam chest 46 will keep the upper wall 22a
of the
cylinder cap 20 heated during operation so as to prevent the condensation
losses due to
chilling of steam entering the steam expansion chamber 44 for avoiding a drop
in mean
effective steam pressure within the cylinder which is most likely to occur
during startup
or when the engine is run intermittently. The engine can be run with a dry
sump to keep
oil away from the piston exhaust openings 14b which are placed in alignment
with legs
201 and 20r and not with slots 20s and 20t. Lubrication is supplied
conventionally by
force feeding from a passage in the crankshaft 30 through the connecting rod
at 28a (Fig.
4) and through oil supply lines 57 to the cylinders. Excess oil and blow-by
combustion
products are stripped off the piston through a channel 56 and return duct 60
as described
in pending application 12/075, 042 which is incorporated herein by reference.
It can be seen that approximately the lower half of the piston is exposed to
ambient temperatures when below rings 20b. Heat transfer from the piston skirt
to the
cylinder wall 12 is minimized by a thermal insulating layer 12c (only part of
which is
shown) covering the entire outside of the engine. Convection losses to air
inside the
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piston below the steam chest 46 are reduced by a pair of semi-circular sheet
metal baffle
plates 20f & 20g extending horizontally across the lower end of the cylinder
cap
assembly with a gap between them that is wide enough for the connecting rod to
pass
through (Fig. 4). The lower part of the piston will equilibrate to a
temperature
intermediate ambient and the mean temperature of the ring area 19 and inside
the
cylinder. Because the top third of the piston is never exposed to ambient
temperatures
and the bottom third is not raised into the steam expansion chamber 44, only
the center
third is able to materially affect condensation losses. Assuming the engine
interior is at
about 180 F following warm up, there should be no condensation loss if the
steam
admission temperature is at least about 350 F. Heating steam to 350 F can be
accomplished without difficulty since combustion exhaust gases are usually in
the range
of about 750 F to 1150 F but can be as high as 1650 F. The temperature of the
piston
skirt will also be raised during operation due to the heat conducted from the
ringed part
19 of the cap which is always the throttle temperature. In any engine where
condensation
losses can occur due to reduced piston skirt temperatures, each piston is
internally
sleeved (Fig. 2A) with a liner sleeve 80 that is bonded inside the piston with
one or more
internal heating chambers 81 between it and the outer piston wall 82, each
connected by
a small metering port 83 to the steam chamber 44 near the piston crown so as
to hold
steam during operation for heating the skirt of the piston from the inside
while also
serving as an auxiliary clearance volume chamber. The ports 83 function in the
same
way as port 47 to be described below. The piston can be ribbed internally and
longitudinal grooves 84 are provided in the outward surface of the sleeve 80
to carry
steam throughout the piston skirt.
Figs. 4 and 4A show a cylinder cap 20 with a groove containing a steam
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collection ring 20c with upwardly opening circumferentially spaced radial
slots 20d that
serve as passages to carry any blow-by steam through a transfer duct 20e that
is at
atmospheric pressure to the condenser to capture escaping steam before it can
be released
into the engine.
Refer now to Fig. 5 wherein the same numerals refer to corresponding parts
already described which shows an alternate form of cylinder cap 20 in
accordance with
the invention that does not have a steam chest 46 but instead has a small
chamber 22c
only in one portion of the cylinder cap below the head of valve 48. In this
embodiment,
instead of a passage 22d that is part of the leg 201, steam at throttle
pressure is admitted
from supply pipe 49 through a pipe 22f to the chamber 22c located below the
head of
valve 48 so that each time the valve 48 is retracted inwardly, the steam is
admitted to the
steam chamber 44 via supply pipe 22f and chamber 22c. Consequently, there is
substantially no heating of the upper wall 22a of the cylinder cap 20 as in
Figs 1-4 but
only incidental heating of the upper wall 22a of the cylinder cap from high
pressure
steam in the chamber 22c. The steam from the steam supply provided through
pipe 22f is
thus kept out of heat transfer relationship with parts of the cylinder cap
other than at the
inlet valve. In this way, overheating is avoided in an installation in which
the engine 10
is run continuously as a battery charging module at a relatively high RPM so
that heating
of the cylinder cap upper wall 22a could produce a loss in efficiency due to
excessive
heat being carried away as the steam is exhausted from the steam chamber 44.
The
chamber 22c can be any size, with or without small steam ducts 61 if needed as
described in copending application 12/387,113 extending from it in the upper
wall 22a to
thereby control heating of the cap 20 to any desired temperature that is
required to
prevent condensation losses on the cap 20, yet without enough heat flux for
permitting
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excessive heat to be carried away from the cap as the steam is exhausted.
Controlled
heating can thus be accomplished by providing the top wall 22a with reduced
heat flux
e.g., by supplying less heat than that furnished by heating the full area of
the cap as in
Figs. 1-4.
Refer now to Fig. 6 which illustrates another form of the invention in which a
pair of series related valves take the place of a single valve 48. In Fig. 6
wherein the
same numerals refer to corresponding parts already described, the valve 48
passes
through a central bored opening within a second valve 48d that is concentric
with valve
48 and has a head that forms a seal on conical valve seat 48e. During
operation of Fig. 6,
the valves 48 and 48d are operated in sequential timed relationship to
precisely control
the cutoff of the steam mass admitted during each cycle with a provision for
changing
the phase of the valves to vary the cutoff continuously from, say, 5% to 50%
of the
power stroke as determined by an engine controller 305 (Fig. 7) more fully
described in
copending U.S. application S.N. 12/387,113 which is incorporated herein by
reference.
Steam at throttle pressure is supplied through the pipe 22f to the chamber 22c
as
described in connection with Fig. 5. Thus, the phase of the camshaft for valve
48d is
advanced or retarded with respect to valve 48 to thereby regulate the cutoff
of steam
through the sequential operation of valves which together permit the admission
of steam
through an intervalve passage into the steam expansion chamber 44 during the
interval
that both admission valves 48 and 48d are open. This enables the steam cutoff
to be
varied throughout operation as determined by an electronic engine management
controller 305 (Fig. 7) through variable cam positioning, namely by providing
a separate
cam shaft for each valve with valve timing control for changing the phase
angle of valve
48d relative to valve 48 (Fig. 6) to control the overlap of the valves. The
valves 48 & 48d
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need not be concentric but can be separate valves connected in series as
described in S.N.
12/387,113 filed April 28, 2009. Concentric valves, however, provide the
advantage of
minimizing the size of the intervalve passage thereby assuring better control
of the
volume of steam admitted, especially at a short cutoff as well as providing a
straight
5 intervalve passage. It will be seen that in both Figs. 5 and 6, because heat
is supplied
from the pipe 22f only to the chamber 22c in one portion of the cylinder cap,
heat flux to
the cylinder cap is less than that produced by jacketing the entire cylinder
cap 20 with
steam from the steam chest as shown in Figs. I and 2. The valves 48 and 48d
are biased
upwardly to their seated position on a valve seat, each by a spring such as
spring 48c.
10 Valve 48 as well as valve 48d can function as pressure relief valves on the
down stroke
of the piston when pressure in chamber 44 exceeds that in 46.
Fig. 1 shows a conventional internal combustion chamber 34 above the upper
face 14a of the piston 14 enclosed at the top of the cylinder by a cylinder
head 35 which
has an inlet valve 36, an exhaust valve 38 and port 37, chambers 39 for
coolant
15 circulation, and a spark plug 40 operating as a four stroke (Otto) cycle
I.C. engine that
burns gasoline or other fuel in the combustion chamber 34 but which can be a
diesel
engine or a two stroke cycle engine, Atkinson or other cycle if desired. The
combustion
chamber 34 is cooled by a coolant at 39 in the head that is also circulated
through a water
jacket 12b (Figs. 1 and 7) of the cylinder 12. It can be seen that the
combustion chamber
34 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
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with the interior of the cylinder 12 through several circumferentially 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 by superheating steam
produced in
a jacket 12b (Fig. 1) which is then supplied to the engine. 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 the 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 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. As the exhaust valve opens, any moisture
on wall
22a will be swept forcefully out of the cylinder thereby avoiding condensation
losses that
may otherwise be caused by residual moisture on or around the top of the cap
20.
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 the water or other working fluid seldom requires
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. 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 (here a
flat plane), the
CA 02708474 2010-06-25
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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. at 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 can be heated externally throughout operation when
advantageous
by the steam chest 46. Low-pressure steam is thus exhausted through line 52
(Fig. 1)
when chamber 44 is 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.
The construction shown in Figs. 1-4 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. 1-7 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. 1-7 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
CA 02708474 2010-06-25
18
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, if desired, 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 which is most likely
during start-up
or intermittent operation. It is of importance to note that the invention
enables low-
pressure steam exhausted through ports 51 and line 52 to be kept away from the
heated
area below cap 20. Any oil or condensate in the steam chest can be removed
through a
drain (not shown).
STEAM EXHAUST
During operation, when the exhaust ports are uncovered by the piston openings
14b acting as an exhaust valve, any moisture on the top of the cap 20 is blown
out of the
cylinder in several directions rather than being left in it to be evaporated
again during the
following power stroke as is the case in a counterflow engine thereby
eliminating
condensation losses. As noted previously, on the down stroke, the remaining
steam in the
cylinder is recompressed in the clearance space to substantially admission
pressure. A
spring loaded steam relief or bypass valve (not shown) can be provided in the
cap 20, if
desired as described in copending application S.N. 12/387,113 to prevent
excessive
pressures in chamber 44 during start up or in case the condenser fails. From
the steam
exhaust manifold 50; the low-pressure steam passes through pipe 52 to a
condenser, next
to a steam generator where it is turned to steam, then through line 49 back to
the high-
CA 02708474 2010-06-25
19
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.
ENGINE MANAGEMENT CONTROL
The engine management control is accomplished by means of a central engine
management control 305 (Fig. 7) as described more fully in prior copending
patent
application Ser. No. 12/387,113 which is incorporated herein by reference for
continuously regulating the various output devices including the IC engine
throttle, the
steam throttle T (Fig. 7) and the cutoff of steam to chamber 44 by means of
valve 48, i.e.,
the point in the cycle at which valve 48 opens and closes or the overlap with
valve 48d
(Fig. 6) for determining the mass of steam admitted each cycle of operation in
order to
maximize the efficiency and reduce the specific fuel consumption of the engine
under the
operating condition being experienced.
Steam admission timing, cutoff control, and steam valve phasing is described
in
copending application S.N. 12/387,113 which is incorporated herein by
reference. The
camshaft 61 can be gear-driven, e.g., using known methods of variable valve
timing for
advancing or retarding the camshaft cam 64 thereby advancing or retarding the
steam
cutoff. Alternatively, each cam 64 of camshaft 61 can be an axially moveable
three-
dimensional cam contoured along its length to provide different cutoff at each
position
set by the computerized electronic motor control 305. Thus, the control 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.
STEAM ENGINE EFFICIENCY AND AVOIDANCE OF ENTHALPY LOSSES
It can be seen that Rankine efficiency is enhanced by the direct conduction of
heat from the burning gas in combustion chamber 34 through the top l4a of the
piston to
CA 02708474 2010-06-25
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 14 and in turn
to the steam
in chamber 44 thereby increasing enthalpy of the steam and enhancing
efficiency. The
5 head of the piston can however be maintained at a safe operating temperature
due to the
large volume of steam passing through the chamber inside the combustion
piston.
From the foregoing description it can be seen that the invention avoids
evacuating low pressure steam throughout the entire stroke of the piston as in
a
counterflow engine that enables moisture to collect in the cylinder. It also
avoids having
10 a dead air space under the piston crown that would interfere with cooling
the combustion
chamber and it eliminates large cylindrical surfaces facing a narrow entry
area in prior
expanders that can condense steam entering the engine. In the present
invention, steam
flows out from the inside of a combustion piston during a short period of time
when
openings in the piston itself are aligned with a ring of cylinder exhaust
ports.
15 Recompression then takes place inside of a combustion piston on the
opposite side of a
wall heated by the hot combustion gasses. Moreover, the cylinder cap 20 can be
heated
throughout its entire area or to any desired degree thereby eliminating
condensation
losses that might otherwise occur without waste caused by excessive cylinder
cap heat.
Communicating with steam chamber 44 through a metering duct 47 as shown in
20 Fig. 1 is an auxiliary clearance volume chamber 45 within the piston crown.
The duct 47
is chosen to regulate steam flow so as to provide a larger effective clearance
volume at
low RPMs but restrict flow increasingly at higher RPMs when less time is
available
during each cycle for chamber 45 to fill. The poor high frequency response of
duct 47
thereby provides a smaller effective clearance volume so as to achieve maximum
CA 02708474 2010-06-25
21
efficiency as described in copending applications 12/075,042 and 12/387,113,
which are
incorporated herein by reference. Therefore, the effective clearance volume
within the
steam expansion chamber 44 is varied depending upon engine speed by the
auxiliary
clearance volume chamber 45 and metering duct 47 (Fig. 1) 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 valve 48 (or in Fig. 6 valves 48 and 48a) at or near b.d.c., 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. An instant later 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 the
admission valve or valves 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 has shown 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
CA 02708474 2010-06-25
22
chamber 44 as well as the capacity for achieving a mean cycle temperature gain
thereby
assuring a higher level of Rankine efficiency.
SUPERHEATER ASSEMBLY
Refer now to Fig. 7. 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 assembly 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 i.e. 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 i.e.
exhaust pipes 141-144 via injectors supplied with air from a blower 148 via
air supply
line 146. It can 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. Combustion exhaust gas entering the superheater
104
CA 02708474 2010-06-25
23
through the exhaust passages 141-144 can be as high as about 900 C (1652 F)
but the
most common range is about 400 C - 600 C. The auxiliary air supply introduced
through
the supplemental air 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 car 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.
Operation of the superheater is described more fully in copending Application
No.
CA 02708474 2010-06-25
24
12/387,113 filed April 28, 2009 and 12/075,042 filed March 7, 2008 which are
incorporated herein by reference.
Fig. 7 shows 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 as described
above
and in application 12/075,042 filed March 7, 2008 and application 12/387,113
filed April
28, 2009. For simplicity and clarity of illustration, the complete engine
coolant circuit
within the engine and the steam generator designated 100 in application
12/075,042 has
not been shown in Fig. 7 but both can, if desired, be constructed as described
therein. In
Fig. 7 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
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
CA 02708474 2010-06-25
used primarily as a back-up battery recharging device to increase the range of
the
vehicle.
Fig. 7 also illustrates how engine cooling and final steam production are
integrated in series by circulating a single working fluid in a closed loop to
serve as an
5 engine coolant as well as the working fluid in the engine. Thus, the fluid
which is heated
first in the combustion chamber cooling jacket 12b surrounding the cylinders
preferably
to form steam by evaporative cooling flows out through pipe 504, then through
a
regenerator 106. To prevent the occurrence of hotspots in the combustion
chamber due to
runaway heating, water in the cooling jacket is agitated preferably by sonic
vibration, by
10 spray cooling or by connecting a pump to circulate coolant in a separate
intrajacket
circuit, i.e., by providing a constant laminar flow of coolant throughout the
jacket 12b
and 39 to maintain nucleate boiling. The steam produced 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
15 the steam expansion chambers 44 of the engine 10 below the pistons 14. By
running the
combustion cooling chambers 39, 12b (Fig. 1) at a high enough temperature to
evaporate
the coolant within the cooling jacket 12a 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 of chamber 500 through a steam duct 504 to a
pressure
20 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 the countercurrent flow heat
exchanger or
regenerator 106 where low-pressure steam exhausted from the steam expansion
CA 02708474 2010-06-25
26
chambers of the engine 10 through line 52 into 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). A
bypass line
(not shown) can also be connected between jacket 12b and line 508 to meter
water to
heater 104 under the control of CEM computer 305 whenever the flow of steam is
insufficient.
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
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 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 in the condenser 520 owing to a substantial
temperature
CA 02708474 2010-06-25
27
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. When evaporative cooling is used, much smaller pumps
511 and
534 are required than in an ordinary automobile cooling system since the flow
rate
required is only that needed to replace the water that is boiled away and 1
lb. of water
which is boiled has about 30 times the cooling effect as in a standard cooling
system.
All references cited above are considered to be disclosed as fully and
completely
as if reproduced herein in their entirety.
Many variations of the invention within the scope of the appended claims will
be
apparent to those skilled in the art once the principles described herein are
understood.
