Note: Descriptions are shown in the official language in which they were submitted.
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Descrlption
Adiabatic Positive Displacement _ chinerY
Back~round of the Invention
Introduction
~ .
There are in general two types of machinery used
to do either work on, or have work done by, the com-
pression or expansion of gases. These two generic
types of machinery are positive displacement and tur-
bine. The positive displacement type includes various
mechanically driven or driving pistons or vane type
rotors. A volume of gas is carried at relatively low
velocity from one volume to a different one, either
larger or smaller depending upon the function of com-
pressor or engine. In the other type of machinery,
turbines, the gas flow through blades occurs at a
velocity of roughly the speed of sound of the gas.
It is well known to those designing such machinery
that the turbines can be made more efficient than
positive displacement machinery. The reason for this
difference in efficiency has frequently been obscure.
A knowledge of the source of this inefficiency will
allow positive displacement machinery to be designed
in a fashion such that the inefficiency or loss is
reduced by a significant factor to a minimal value.
There is, of course, the well-recognized, additional
loss of energy in positive displacement machinery due
to the friction between whatever is the displacer,
piston or vanes, and the walls of the chamber. The
turbine in turn avoids this inefficiency but has others
such as the friction of aerodynamic flow at velocities
near the sound speed.
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Heat Exchanqe and Total Energy Loss
Frictional loss between sliding parts is impor-
tant, but not usually the principal energy loss in
the system. However, I will focus on one property of
positive displacement machinery that does cause a
major inefficiency and that is not well understood.
This is the heat exchanse between the gases being
compressed or expanded and the walls of the positive
displacement volume. This heat exchange iis usually
accepted as fundamental. Instead, I claim it can be
significantly reduced.
Heat Exchange with the Walls
Let us consider first compressors, although these
comments can be equally applied to expansion engines
with an inversion of terms. If a gas is adiabatically
compressed, it becomes both hotter as a function of
compression as well as increased in pressureO The
increase in temperature and pressure follow the well
known relations of the adiabatic law. In some cases,
as in an air compressor, the additional temperature
created in the gas is later rejected as a nuisance,
although a significant fraction, even a major frac-
tion, of the useful work may be wasted in the rejec-
tion of this heat. In the peculiar case of an air
compressor, where this heat is rejected, it is more
efficient to reject this heat as early in the cycle
as possible so that less work is done achieving a
desired volume of cold compressed gas. (Isothermal
machinery is described in Canadian patent application
No. 411,265 entitled "Isothermal Positive Displacement
Machinery" and filed September 13, 1982 by Colgate
Thermodynamics Co.) In other cases where a compressor
is used, as in a Rankine cycle heat pump or compres-
sion cycle of various internal combustion engines,
this departure from an adiabatic compression due to
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heat exchange of the working fluid~ i.e. gas, with
the walls of the compressor is a major disadvantage
and inefficiency of the system. A point of this
invention is that by proper design of the input and
output ports of adiabatic positive displacement
machinery this heat exchange can be reduced to a small
value.
The mechanism for this heat loss is turbulent
motion of the working fluid making contact with the
walls during compression or expansion. There are two
parts to this heat exchange: (l) the heat exchange
between the gas and the wall if the wall were held
isothermal, and (2) the heat impedance of the wall
itself. It turns out that the heat impedance of the
lS wall is such that the wall acts as a time lag averag-
in~ reservoir coming to a temperature equal to the
mean temperature of the gas at a delayed phase of the
stroke. The time phase lag as well as the magnitude
of heat exchange are both detrimental to adiabatic
efficiency.
Thermal Skin Depth
One can calculate the heat mass of the wall ~ur-
ing the transient contact with the gas by calculatin~
the thermal skin depth within the time o~ heat contact.
The thermal skin depth, d, of penetration of heat (or
cold) within the given time t is expressed mathemati-
cally as
d = [(K/Cv t] J
where Cv is the specific heat of the wall material, K
is thermal conductivity, and t is the timeO (K/CV)
is oten called the diffusion coefficient. For typical
materials where Cv is 1 calorie cm 3 deg 1, and the
~ime = 10 2 sec (for a stroke at 3000 ~lP) or longer,
the skin depth will vary between 3 x 10 3 cm for a
plastic with K = 10 3 cal cm 3 deg 1 at the highest
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speed to 3 x lO 2 cm for a metal and a large slow
piston. Even the smallest skin depth corresponds to
a heat mass equivalent to several centimeters of air
or freon at atmospheric pressure. Therefore the heat
S mass of the wall in contact with the gas will be com-
parable to or larger than the heat mass of the gas.
It is usual in engineering practice to neglect this
skin depth factor and assume that the wall takes on a
temperature which is the time average of the heat
flow from the gas. In this case the primary factor
in determining heat loss is the theoretical heat
exchange of the gas with an assumed isothermal wall
almost independent of wall properties. Later I will
show the importance of the time dependent phase lag
of the heat flow. First I will demonstrate the skin
depth effect. We assume that the walls of the chamber
will be smooth and then the heat loss will be governed
by the turbulent flow exchange with a smooth wall.
Description of the Drawings
Fig. l is a diagram depicting heat transfer across
a barrier;
Fig. 2 is a PV diagram illustrating various
thermal cycles;
FigO 3 is an end cross-sectional Yiew of an
articulated vane compressor-expander;
Fig. 4 is a diagram depicting the temperature
drop in the gas in the machine of Fig. 3;
Fig. 5 is a schematic drawing of a Brayton cycle
heat pump using the vane machines of Fig. 3;
Fig. 6 is a side cross-sectional view of a com-
pressor;
Fig. 7A is a top cross-sectional view through an
induction passage showing radial vanes -- such radial
vanes may be used but are not preferred;
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Fig. 7B is a top cross-sectional view through
the induction passage in the compressor shown in Fig.
6 -- these vanes are oblique to the radial and induce
the desired axial vortex flow that suppresses the
formation of an annular vortex;
Figs. 8A, 8B and 8C illustrate the annular vortex
that forms and persists at bottom stroke, mid stroke
and top stroke, respectively, when gas is inducted
into a piston machine with little or no circumferential
component of velocity -- such would be the case with
radial vanes like the ones shown in Fig. 7A;
Figs. 9A and 9B illustrate schematically the
formation and expansion as induction progresses of an
axial vortex flow as a cylinder expands -- such a
flow is induced by the oblique vanes shown in Fig.
7B;
Fig. 10 is a cross-sectional view of a two-stroke
diesel engine embodying the present invention;
Fig. 11 is a top view in generally schematic
form illustrating the path of gas flow through the
engine shown in Fig. 10;
Fig. 12 is a fragmentary cross-sectional view of
the head of a combustion cylinder that can be used in
lieu of the head shown in Fig. 10;
Fig. 13 is a partial cross-sectional view taken
along the lines 13-13 of Fig. 12 and in the direction
of the arrows;
Fig. 14 is a diagram showing the timing of the
engine shown in Figs. 10 and 11;
Fig. 15 is an end cross-sectional view of a four-
stroke Otto cycle engine embodying the present
invention; and
Fig. 16 is a top cross-sectional view of the
engine shown in Fig. 15 taken at the induction passage,
as indicated by the lines 16-15 in Fig. 15.
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Explanation of Diffusive Heat Flow
In Figure 1 I show the classic solution of the
diffusion of heat from one reservoir 1 into a second
reservoir 2. Let us assume that 1 is hotter at T1
and is a turbulent gas with essentially infinite abil-
ity to transport heat up to a barrier 3. The heat
diffuses into, or out of, region 2 with a dif~usivity
K/Cv. Then the distribution of heat or temperature,
T, as a function of depth, x, follows a sequence of
"error function" solutions in which
T = T2 ~ (Tl - T2) exp(-x2/d2)
or T = T2 + (Tl w T2) e( x /d )
where as before
d = [(K/C t]l/2
The distance d is t~e centroid of the depth of penetra-
tion of the thermal wave. The three curves labeled
dl, d2, d3 are the temperature profiles of times tl,
t2, t3 where tl t2 t3 with characteristic skin depths
dl is less than d2 is less than d3. If Tl is time
dependent as it would be in a cylinder with alterna-
tively hot or cold gases, then the actual distribution
of temperature should be a simple addition of such
solutions. In this sense "cold", i.e. Tl is less
than T2, can penetrate into the wall just ~s well as
hot, Tl is greater than T2O The skin depth îs just
the characteristic a~eraging depth of each temperature
variation in a time t. The heat mass described by
each curve is H = (Tl - T2)CV and hence the longer
the time the heat has to "soak" inr the ~reater the
heat transferred. Typical diffusivities and skin
depth heat masses are shown in Table 1 for various
materials. A frequency of 3000 RPM is chosen as an
example and the skin depth heat mass is compared to
8:1 compressed combustion gases typical of an Otto
cycle engine.
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TABLE 1
Diffusivity, skin depth, heat mass of various materials
assume 3000 RPMI t = 1/(2 f) = 0.01 sec.
Thermal lleat Diffusivity Heat mass of
5Conductivity Capacity K/CV skin depth
wattS/cm2Cal cm~3 cm2 sec-l ~V(Dt)L/2
cal cm~2
Carbon 0.5 0.81 0.13 0.0164
Steel
Stainless 0.14 0.81 0.036 0.0087
Steel
Nickle- 0.11 0.81 0.028 0.0076
Chrome
Phosphorus 2.2 0.84 0.55 0.035
Bronze
Berylilium 0.8 0.84 0.20 0.021
Copper
Aluminum 1.6 0.58 0.57 0.025
Alloy
Carbon 0.28 0.3 0.2 0.0075
Coke
Aluminum 0.30 0.8 0.08 0.013
Oxide
Ceramic
Silicon 0.016 0.8 0.004 0.003
Dioxide
Fused
Heat ca~acity of air plus fuel 8 fold compressed = 5 x 10-3
cal cm~
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Turbulent Heat Exchange with a Smooth Surface
If a gas flows in a smooth-wall pipe, then the
properties of turbulent fluid heat exchange are such
that the gas will reach thermal equilibrium with the
wall after moving roughly 50 pipe diameters (American
~andbook of Physics, 1963). This is also the viscous
slowing down length, or the length in which kinetic
energy is dissipated. The quantity "50 pipe diameters"
is determined by the peculiar properties of the laminar
sub-layer. This is the boundary layer between turbu-
lent fluid flow and smooth pipe wall. In the case of
the cylinder or other compression volume the appropri-
ate consideration is the distance the fluid (or gas)
travels in contact with the wall during the time of a
stroke~ If the gas enters from a valve with a high
velocity relative to the chamber, then the gas will
circulate many times within the compression chamber
during the time of a compression or expansion stroke.
The number of cycles of circulation can be roughly
estimated by the ratio of the velocity of the gases
entering through the input valve to the velocity of
the piston. The average ratio of the valve area to
piston area is frequently about 20 to 1 (Taylor, 1966),
so that gases entering the cylinder have velocities
between 10 to 20 times that of the piston velocity.
In general the gases enter the chamber non-symetrically
with respect to the compression volume so that the
turbulence generated by the flow will be greater than
that induced in a normal pipe flow of a fluid moving
through a pipe. ThereEore the heat exchange with the
wall will be greater when the turbulence is greater~
We expect roughly e-fold of heat exchange within
roughly 10 circulation times because the gas flowing
by corners will be more ~urbulent than straight pipe
flow. Therefore the typical piston with restricted
inlet valves will allow heat exchange of the gas with
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the wall of roughly half the differential heat of the
gas during the time of compression or expansion stroke.
Since the differential temperature of the wall rela-
tive to the ~as is roughly 1/2 the total temperature
difference, then roughly 1/4 of the heat is lost to
the wall. It is this large heat exchange which
accounts for the primary inefficiency of such gas
handling machines. The only way to avoid this heat
loss is to allow the gases to enter the compression
volume with low velocities. Then the distance the
gas moves during a stroke is small (measured in diam-
eters) and the heat exchange will be small. If the
flow velocity of the entering gas carefully matches
the velocity of the piston or other compression mem-
bers, then we expect a weakly turbulent boundary layer,i.e. not perfect laminar flow but instead a low turbu-
lence. The near absence of turbulence I call near-
laminar flow and hence the crucial design is to create
near-laminar flow of the input gas to the compression
or expansion cycle. If the flow is to be near-laminar,
at the piston velocity, then the inlet port area must
be close to the full piston area. Or similarly in an
expansion engine the inlet ports must again be equal
to the piston area. This also applies to rotating
vane machinery.
The Inefficiency Due to the Exchange of Heat
of the Gas with the Wall in an Adiabatic Cvcle
Suppose a gas initially at temperature Tl is
compressed such that its final temperature would be
T3 if it were a perfect adiabatic compression but
instead is held isothermally at an intermediate tem-
peratuee T2 durin~ the latter part of compression.
Then Tl is less than T2 is less than T3, and then the
heat energy in the gas after it leaves the piston
will be less than it would be by the ratio T2/T3.
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(The mass of the gas is conserved). Therefore the
înefficiency factor or the heat loss is just the dif-
ference (T3 - T2) divided by the heat that would have
been in the gas (T3 - Tl). Depending upon the cooling
of the cylinder walls and other factors T2 might be
only half way between Tl and T3~ and therefore com-
pression machinery would be 50~ efficient in following
an adiabatic compressionO The temperature T~ that
the wall reaches will be a complicated function of
1~ the heat exchange process and the cooling of the walls.
In general the gas will not come into equilibrium at
every point in the stroke, and so only an approximation
to this heat loss will actually occur. However, the
fact that a simple calculation indicates that up to
50% of the theoretical maximum heat can be exchanged
is sufficient reason to try to design machinery where
one avoids this heat short circuit and its attendant
loss in efficiency.
If the wall remained isothermal at temperature
T2, then this heat loss to the walls would be an actual
advantage in a compressor as, for example, a refrigera-
tion cycle or normal air compressor. However, the
heat exchange of the gas to the wall is moxe compli-
cated than this. If the gas can lose heat to the
wall in part of the cycle it can also gain heat from
the wall in another part of the cycle if the wall is
hotter than the gas. The wall will be hotter than
the gas for a transient time due to the skin depth
effect. This latter effect of heating the gas from
the wall is particularly harmful to the efficiency of
the compressor because the heating of the ~as occurs
at its induction when the wall is hotter than the
inlet gas. The gas is then compressed with higher
heat than the ideal adiabatic cycle, and hence more
work is required than would be required for the idea-
lized cycle. Thus the heat is exchanged with a harmful
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phase lag. Let us illustrate these ideal cycles with
and without heat exchange with the wall, figure 2.
The gas is drawn into the cylinder during the
induction stroke starting at temperature To along the
constant pressure PO to the volume, VO. In the ideal
cycle it starts compression at volume, VO along the
pure adiabatic curve 1, reaching the final reservoir
pressure Pl at volume ~1 and temperature Tl. Several
possibilities due to heating the gas by the wall exist.
(1) If the gas heated by + Tdiff only during
induction, then the pressure volume relation will
remain the sameO That is, since the gas is only heated
by the walls during induction and not during compres-
sion, by assumption, the compression will be adiabatic
and therefore will arrive at the same state Vlr Pl,
but at a higher temperature T = (Tdiff + To)/To x Tl.
The excess heat will be later rejected, therefore
requiring more work to deliver the same mass of gas.
(2) Heat can be added after the start of
compression and the gas will follow the curve 2,
steeper than the pure adiabatic one. The gas tempera-
ture is then likely to exceed the wall temperature,
transferring heat from ~he gas back to the wall and
the curve will bend over, curve 3, less steep than
the adiabatic curve 1. The work required will be
greater. Curve 4 is more realistic in that wall
cooling of the compressed gas at the end of the cycle
may actually reduce the final gas temperature, T4 at
V~, below Tl at Vl of the adiabatic case, but the net
work still exceeds the adiabatic case.
(33 The wall can be cooled perfectly and
retained at the temperature To, the gas can exchange
heat with the wall perfectly and then the compression
is isothermal along curve 5. This is the minimum
work cycle to obtain cold gas at the final temperature
T5 = To~ It usually cannot be achieved in practice,
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again because (1) the skin depth argument that isolates
the interior from the exterior on a transient basis,
and (2) turbulent heat exchange is only partially
effective in a normal cylinder and piston.
Summar~ of Heat Loss and Adiabatic Cycle
The heat exchange occurs because of turbulent
flow in the induction gas. The maximum gas mass or
minimum temperature To is maintained during induction
only if either the walls are retained at temperature
To or induction is near-laminar flow. Durîng compres-
sion the same argument applies. However the thermal
skin depth argument says that if the wall is thick
compared to the skin depth, it will average the heat
flow on the outside, but inside it will alternately
be hot and then cold in a thin layer. If the gas is
turbulent, this alternately hot and cold heat reser-
voir will cause heating of the induction air at the
worst time, causing the compressed gas to reach a
hotter temperature T3 that in turn heats the gas still
further and requires still more work~ and so forth,
until the higher average temperature of the walls
allows the heat to be carried away. This is an ineffi-
cient compressor. It is better to reduce the heat
exchange between the gas and walls by decreasing the
turbulence and having near-laminar flow incluction as
well as compression.
The transient heat exchange due to partial turbu-
lence and thermal skin depth is deleterious to all
positive displacement heat machinery. As a useful
measure Taylor (1966) ascribes about 30% efficiency
loss to heat loss in a gasoline engine and up to 50%
heat loss in a diesel engine. In other words a gaso-
line engine could be 45% efficient instead o 30% and
a diesel could be 70% rather than 35% to 40%. These
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are large potential gains and therefore warrant some
degree of complexity to achieve them~
Summary of the Invention
Laminar_~ea~ Exchan~
If the fluid exchanges heat with the wall in
something like 10 diameters of motion, assuming the
volume has corners, it means that the velocity of the
incomins fluid cannot be very much greater than the
velocity of the piston or other moving boundary of
the confinement volume (e.g., rotating vane). Since
~hese velocities in typical machinery are usually
less than 1/10 of the speed of sound of the gas, it
means that the differential pressures at the inlet
ports can be no greater than roughly 1% of the gas
pressure. This in turn means that the ports must be
designed with ~an] area about equal to that of the
piston. Valves that open by the gas pressure, like
reed valves, will necessarily cause a h19h enough
velocity of the inlet gases to give rise to hi~h tur-
bulence levels and hence large heat exchange losses,
but a sleeve type intake valve will serve well. The
exhaust port, on the other hand, need not be so large,
and indeed this can be a reed operated valve, because
the gases leaving the cylinder do not create turbu-
lence within the cylinder during the exit process.
~s a consequence it is possible to make, for example,
a laminar-ported piston compressor but only with the
relative complication of a sleeve type inlet valve
that exposes a side wall area. This area should be
almost equal to the full area of the head of the
cylinder, but substantial reductions in heat loss are
provided with smaller induction ports, say about one-
half the cylinder head area.
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Vane type machinery, on the other hand, can be
designed such that the inlet port area is as large as
the entire compression or expansion volume cross sec-
tion, provided the vanes do not ride on the outside
wall of the compression volume. Otherwise the port
entrance would have to be somewhat restricted by the
necessary webbing to support the vane as it goes by.
I therefore provide specially designed ported
engines or compressors in which the inlets are roughly
equal to the entire cross sectional area of the com-
pression volume and special care is taken so that the
inlet gases enter with about the same velocity as the
moving boundary of the confinement volume. In this
fashion the gas flow will be near-laminar within the
confinement volume during compression and expansion,
and the heat loss to the wall will be significantly
reduced. In many heat engines this should allow an
improvement in fuel efficiencY of uP to a factor of
t
Laminar and Turbulent Flow
Turbulence is generated in fluid flow when two
conditions are met: (1) the viscous dissipation of a
kinetic energy of the average flow field is small, or
equivalently the Reynolds number of the flow pattern
is large; (2) the gradient of the velocity distribution
is no a constant; that is, there exist finite higher
derivatives than the first derivative of the velocity
as a function of distance perpendicular to the mean
flow. Therefore, a uniform shear in the flow is not
sufficient to initiate turbulence.
In practical terms turbulence enhances the fric-
tion (and heat transfer) of fluid flow in contact
with a rigid surface. As one progresses from a smooth
- surface into the fluid, the flow immediately adjacent
to the wall is laminar because the dimension i5 SO
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small that fluid friction caused by viscosity is
greater than the turbulent friction. At a critical
distance into the flow where the Reynolds number (mea-
sured perpendicular to the wall) is greater than 100,
the flow becomes turbulent, first with small eddies
because there is only room for small eddies and then
progressively larger eddies as one progresses further
into the fluid. The progression of eddy sizes as one
proceeds away from the surface into the fluid is called
the "logarithmic profile." As one proceeds further
downstream, for instance along a wing of an airplane,
the turbulent profile extends further into the fluid.
This depth of penetration for a smooth surface is a
small fraction, about l/lOth to 1/2~th, of the down-
stream distance. Hence the flow travels a relatively
long way before exciting turbulence throughout the
fluid. This is because a smaller eddy close to the
wall must excite larger eddies further into the fluid,
and so on. On the other hand if the wall is very
rough with perturbations or projections that are large,
then the turbulence will be excited very rapidly and
eddies of the order of the roughness or projection
size will be formed immediately. An airplane wing is
made smooth so that a ratio of lift to drag in the
range of 10:1 to 20Ol is achieved, but if a spoiler
(vertical projecting flap) is used, the turbulence
created is large, and the lift to drag ratlo falls to
2:1 or 3:1. On the other hand if the flow were per-
fectly laminar, a lift to drag ratio of greater than
100 would be possible. Hence on a relative scale the
smooth wall with a weakly turbulent boundary layer
acts as if it were "near-laminar" in drag characteris-
tics, as opposed to extreme turbulence as would occ~r
with a spoiler.
In the case of rotational flow in a cylinder,
the flow is in contact with a smooth wall with no
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corners and hence it is near-laminar. An azimuthal
vortex, on the other hand, will hav~ the flow deflected
by sharp corners and hence will be more turbulent.
This is why the axial vortex is called "near-laminar"
whereas the azimuthal vortex is fully turbulent.
Finally, the shear in the velocity distribution as a
function of radius within the axial vortex does not
induce turbulence because the gradient of the velocity
with radius is a constant (condition 2). Only the
contact or friction with the wall induces turbulence.
In this discussion of near-laminar flow along a smooth
surface it is presupposed that the flow immediately
upstream from that surface is itself near-laminar and
of a velocity and direction not greatly different
from the velocity along the surface in question.
~eneral Description of Adiabatic
Positiv~ Displacement Machines
In accordance with the present invention, the
efficiency of positive displacement machines (both
piston-cylinder and vane compressors an~ expanders)
is substantially improved by introducing the gas into
the compression or expansion chamber through an inlet
passage (or several inlet passages) that is shaped
and sized to provide near-laminar flow of the gas
into the chamber, thereby to reduce substantially
heat flow to and from the walls of the chamber. In
the case of piston-cylinder machines, near-laminar
flow is attained by an inlet passage having an area
of from about one-half of to about equal to the area
of the piston, preferably a passage or passages that
open at an inlet port that extends 360 around the
cylinder and is opened and closed by a sleeve valve.
The passage should be a plenum or two or more volutes
that are arranged to introduce the gas into the cylin-
der with a substantial component of velocity tangential
lZ~29C~
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to the cylinder and thereby induce axial vortex flow
and inhibit the formation of a radial vortex and con-
sequent high turbulence and heat exchange. In the
case of vane machines both the induction passage and
the discharge passage have cross-sectional areas along
their lengths such that near-laminar flow at a velocity
substantially matching that of the rotor vanes is
maintained therein, thereby to reduce heat flow between
the gas and the rotor vanes and the casing walls of
the machine. The casing walls of a vane machine are
preferably made of a material having a low thermal
conductivity to reduce heat flow within the walls and
minimize the heat short circuit of the walls.
The present invention includes, but is not limited
to, the following machines:
Four-Stroke Otto Cycle Enqine
The inlet passage extends 360 around the top of
the cylinder as a plenum and is opened and closed by
a sleeve valve operated by the engine crankshaft or
an overhead camshaft. Fuel is introduced into the
cylinder generally along the axis so that it is gen-
erally localized for combustion in a region spaced
apart from the cylinder walls. The plenum has vanes
oriented obliquely to the tangential to induce axial
vortex flow in the chamber. The axial vortex flow
promotes combustion by centrifuging drople~s of
unburned fuel outwardly from the axis where air
undepleted of oxygen is available to support combus-
tion of the fuel.
Two-Stroke_Diesel Engine
This engine has a supercharging piston-cylinder,
a combustion piston-cylinder, and an exhaust piston-
cylinder having, respectively, compression ratios in
the following ranges: supercharging -3:1 t:o 8:1;
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compression - 3:1 to 4:1; exhaust - 6:1 to 9:1. Air
is inducted into the supercharging cylinder through a
360 inlet port, preferab:Ly at the top, that is opened
and closed by a sleeve va:Lve and receives air from a
plenum or volutes that induce a circumferential veloc~
ity component for formation of an axial vortex and
near-laminar flow in the cylinderO The supercharged
air is conducted to an insulated storage chamber which
holds the air for subsequent quasi-static displacement
and scavenging of the combustion cylinder. The volume
of the storage chamber should be in the range of from
about one to about six times the displacement volume
of the combustion cylinder. The combustion cylinder
has a 360 inlet port at the bottom of the piston
stroke that receives supercharged air from the storage
chamber from a volute that induces an axial vortex
flow in the combustion cylinder. Preferably, at peak
compression the radius and stroke of the combustion
cylinder are roughly equal to provide a large clear-
ance volume for minimizing heat loss. The head ofthe combustion cylinder is smooth, and the exhaust
valve is substantially coaxial with the cylinder axis.
Fuel is injected substantially along the axis of the
combustion cylinder so that the axial vortex flow is
substantially undisturbed and combustion is enhanced
by the centrifuging of fuel droplets to regions where
the air has not been depleted of oxygen. The exhaust
valve opening from the combustion cylinder should be
located generally midway between the axis and the
wall of the cylinder, where it provides more complete
scavenging. In the case of a coaxial tubular exhaust
valve~ the head of the combustion cylinder is cooled
to provide a high heat transfer for cooling of the
valveO The exhaust gas is conducted from the combus-
tion cylinder to the exhaust cylinder through a ther-
mally insulated, smooth-walled passage to a volute
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that induces an axial vortex, near-laminar flow in
the exhaust cylinder. The exhaust valve of the exhaust
cylinder is a sleeve valve, is located generally at
half radius and has a width equal to about half the
cylinder radius.
Gas comPressor
Air is inducted into the cylinder from a plenum
having vanes oblique to the radial through a 360
induction port at the top of the cylinder that is
opened and closed by a sleeve valve. The inducted
air exchanges little heat with the cylinder walls due
to the inducement of a near-laminar axial vortex flow.
Articulated Vane Compressor or Expander
Both the induction and discharge passages leading
to and from, respectively, ~he compression-expansion
zone have cross-sectional areas such that near-laminar
flow at a velocity substantially matching that of the
rotor vanes is maintained in the passages~ A highly
efficient Brayton cycle heat pump utilizes appropri-
ately sized articulated vane machines embodying thepresent invention. The housing and shaft coupling
should be insulated to minimize the heat short circuit.
Turbulence for Combustion
It should be recognized that in internal combus-
tion engines turbulence is very often purposefullyinduced in order to "scavenge" the combustion gases
and particularly in the case of a gasoline engine to
induce more thorough mixing of the fuel-air mixture
near the cold walls with hot burning gases in the
interior to promote complete combustion. These
requirements obviously conflict with near-laminar
flow.
12~2~C)
20-
On the other hand the positive control of the
gas motion in both diesel and Otto cycle engines offers
the possibility of designing the fuel injection system
for diesel and fuel-injected Otto cycle engines so
that the cooled boundary layer in contact with the
wall contains little fuel, i.e. is very lean. The
combustion zone is then isolated from the cold outer
cylinder wall and is exposed only to the hotter piston
crown and head. The un-burnt fuel problem is then
reduced.
Compression and Exeansion
There is also the additional conflict of the
large difference between the mean temperature of the
gases during compression and expansion. The tempera-
ture of the gases during compression is very muchless than that during expansion because of combustion,
so the heat loss from the hot gases during expansion
tends to heat the walls to a higher mean temperature
than would occur during compression. Thus the compres-
sion will be hotter than an adiabatic compression andhence require more energy than necessary. Some of
this energy is recovered during expansion, but the
net effect is inefficiency. It is for this reason
that there is a distinct advantage from the standpoint
of efficiency to separate compression machinery from
the expansion machinery. This is done in yas turbines,
but sincP the blades of the expansion turbine must
come to an equilibrium temperature with the highest
temperature gases after combustion, the stresses due
to the required high velocity of the blades severely
limits the peak temperature and hence limits the Carnot
efficiency. One is thus limited with current machinery
by either turbulent heat exchange of the gas with the
walls, or by the limiting temperature of turbine
blades.
-21-
Heat Pumps
There is a special advantage of near-laminar
flow compressors and expanders in Brayton cycle heat
pumps. There are three general types of heat pumps:
(1) An isothermal cycle or Stirling cycle covered in
my concurrently filed patent application referred to
above. (2) A Rankine Cycle that uses a special
refrigerant that is compressed as a gas, gives up its
heat in a condenser and becomes a liquid. The liquid
is expanded to a gas in a cold heat exchanger where
heat is added. (3) ~ Brayton cycle in which a gas is
adiabatically compressed, heat is extracted in a heat
exchanger then the residual energy is extracted in a
heat expansion engine and finally the heat is added
to the exhaust gas in a second or cold exchanger. In
a Rankine cycle the energy corresponding to the fluid
expanding through the expansion orifice (pressure x
volume of fluid) is wasted, but since the volume is
small because of the relatively high density of fluid,
this wasted energy is small. On the other hand the
restriction of refrigerant properties requires that a
relatively large compression ratio be used so that
the temperature ratios are large enough to cover useful
extremes encountered in average climates. When the
compression cycle is included, say 80~ efi-icient, the
result is an average "coefficient of performance"
(COP) of about 2 to 2.5; the ideal COP is T3/(T3 -
T2). Hence for a typical 30C temperature difference,
and the absolute temperature of 3003K, the theoretical
maximum COP should be 10, not the relatively poor
value of 2Ø In order to approach this higher value
requires that one use a more efficient compressor as
well as avoid the limitations of refrigerants. If
one uses a Brayton cycle, one must now add an expan-
sion engine, and the efficiency of this expansionengine becomes crucial. In such a cycle the compressor
12~Z31~
-22-
does an amount of work (T3 - T2) x (a unit heat mass
of gas) and then an amount of heat that is extracted
in a heat exchanger is this same value (T3 - T2).
This unit heat mass of gas is reduced in volume from
Vol3 to Vol~ in the ratio Vol~/Vol3 = T2/T3. When
the smaller Vol2 is expanded back down to atmospheric
pressure, i.e. the same pressure ratio, it will cool
by nearly the same temperature di~ference (T3 - T2),
but the volume will be smaller and hence the work
done in the engine will be less than that done by the
compressor by the ratio Vol2/Vol3 = T2/T3. If th;s
work is fed back to the compressor, the net work that
must be supplied from the outside is l-(T2/T3) = ~T3 -
T2)/T3 or the inverse of the theoretical maximum ~OP.
This COP of 10 implies that the expander is doing 90%
of the work done by the compressor and therefore the
circulating power is ten times the power being sup-
plie~ from the outside, i.e. the make up power = 10%.
Hence if both the compression engine and the expansion
engine each wastes 5% of its energy, i.e. each is 95%
efficient, then the outside source must supply this
additional lost energy, which doubles the energy that
must be supplied. Hence the COP will decrease from
the theoretical maximum of 10 to 5, or a loss of effi-
ciency of a factor of 2 simply because each engine isonly 95~ efficient rather than 100% eff;cient.
Thus we can see how sensitive heat pump machinery
is to the efficiency of the compression and expansion
engines. Hence there is a major motivation for Brayton
cycle pump machinery to be made as highly efficient
as possible.
A ten percent loss in an internal combustion
engine is not as serious, but then the temperature
differences are much greater so that the loss of effi-
ciency for a given turbulent heat exchange rate is
~Lh~(~L~90
-23-
significantly larger. As a consequence the eficiency
loss is large enough to be significant.
Summary
I therefore provide positive displacement machin
ery -- compressors, expansion engines, internal combus-
tion engines and heat pumps -- in which the gases are
introduced into the positive displacement volume in
such a fashion to be near-laminar, that is, with resid-
ual circulation or eddy velocitîes that are smaller
than or comparable to the positive displacement veloc-
ity of the moving chamber boundaries.
Description of Exemplary Embodiments
Articulated Vane Compressor-ExPander
One embodiment of the present invention is the
articulated vane air pump or expansion engine shown
in figure 3. Articulated vane air pumps are used
extensively in conjunction with automobile gas engines
in order to add compressed air to the exhaust stream
and reduce unburned gases. There is a particular
advantage for using these articulated vane compressors
or expanders (AVC) for heat pumps because of the very
low friction of the moving parts -- the vanes ride on
central shaft bearings as opposed to riding upon an
outside race where the friction is much larger. A
considerable patent literature exists for articulated
vane pumps, such as for automobile emission control,
but this prior art is not directed to the efficiency
of adiabatic compression. Selected citations to the
principal technology are listed at the end of this
specification.
-24-
Inside a housing 72 with inlet and outlet chambers
61 and 62, a set of vanes 63, 64, 65, are rotated
about a stationary axle 66 by a rotating drum 67.
The vanes slide radially through seals 68, 69,70, in
the drum and are sealed by clearances 71 with a con-
centric housing 72 in the compression region and by
the web 73 that separates the inlet chamber 61 from
the outlet chamber 62. The clearance 71 between vanes
and compression housing 72 and between drum 67 and
web 73 must be kept small in order to prevent leakage
of gases, yet still prevent contact since the surfaces
are not lubricated, but instead remain bare to prevent
friction. In the usual forms of state-of-the-art
AVC's the inlet and outlet chambers are of relatively
arbitrary configuration. In this circumstance of a
relatively large chamber, the air flow in and out is
relatively stationary, compared to the movement of
the vanes 63, 64, 65. Hence large turbulent eddies
will be generated by the vanes. In the inlet chamber
these eddies will be captured and the circulation in
the compression chamber will cause increased heat
exchange with the wall. In the delivery chamber the
corresponding eddies will just cause waste heat and
energy. In this embodiment, and according to the
invention, the inlet and outlet chambers are shaped
and dimensioned (dimensions 74 and 75) such that the
flow velocity of the gas just matches the rotation
rate of the drum 67. This dimension 74 for the inlet
61 is equal to the mean vane extension in the compres-
sion volume so that a given rotation of drum 67 andthe vanes will displace gas in the chamber 61 at the
same velocity as the drum 67. In the outlet chamber
62, the width 75 is just that of the inlet dimension
74 divided by the compression ratio, 1.336:1 in this
case of a three-bladed pump as described next.
lZ~ ~ Z~r`
-25-
The rotor vanes 63, 64, 65, are shown when the
compression volume behind vane 65 is just being
released to the outlet 62 when the clearance 71 opens.
The three vane design has an angle of 120 between
the vanes. A design with two, four, five, etc.,
equally spaced vanes can easily be made. The number
of blades determines the compression ratio, which is
simply the volume ratio trapped between the blades
from position I where compression begins to the com-
pressed position II 60 degrees later. The resultingcompression ratio for three blades is a near optimum
value for heat pumps because the compression ratio is
1.336:1 and temperature r~tio is 1.123:1, or Tdiff =
37C and the ideal COP = 8:1. If the compressor and
expander are 95% efficient then the practical COP
will be 4.
In the angular distance of 60 degrees from I to
II the gas is compressed to the pressure at which it
will be released. The blade position 76, indicated
by the phantom lines, corresponds to when the inlet
volume is just being closed. The gas in contact with
each blade at position 76 is ultimately moved to posi-
tion 77 in contact with the outer wall . In the pro-
cess it is heated by compression (cooled if the flow
is reversed for expansion) and hence will heat the
wall. How~ver, in the ideal case the gas and the
wall will come to the same temperature for each posi-
tion corresponding to the compression as the gas is
swept from I to II. Hence if there is no heat con-
duction in the outer wall material, then heat transferwill be minimal. In this case it is, according to
this invention, desirable that the outer wall be made
of low enough conductivity material~ such as stainless
steel or a plastic-coated metal, and be thin to reduce
this heat conduction backwards in the wall of the
housing. The drum 67 and blades 63, 64, 65, however,
12~4Z~30
~26-
rotate from hot region to cold region and hot etc.
The heat transfer from and to the gas is the skin
depth heat diffusion discussed earlier. By making
the flow near-laminar, then the diffusivity of the
gas is less, and the temperature drop occurs in the
gas as depicted in figure 4.
Region 1 is hot at a temperature Tl that is
higher than temperature T2. The mean wall temperature
T3 is the boundary temperature w;th Tl gr~eater than
T3 greater than T2; and the temperature drop (T3 -T2)
by diffusivity into the wall is shown with the skin
depth d2, which is small compared to the laminar gas
where (Tl -T3) is large with the skin depth dl. How-
ever, the density of the gas is very much less than
that of the wall in the ratio of 3 x 10 4 and so the
heat mass is small and the heat lost or exchanged is
small compared to the case in figure 1 where the gas
was assumed turbulent and the heat flux to the bound-
ary 3 was considered large.
~eat Pump
Figure 5 shows schematically a Brayton cycle
refrigeration heat pump ~hat utilizes the laminar
flow AVC air pumps shown in Figure 3. The compressor
81 compresses and heats air entering the intake 83
and discharges it at the outlet 84. The hot compressed
air goes to a standard heat exchanger B5, is cooled
and goes to an expander 86 that is built like the
compressor 81, but smaller in volume flow by the ratio
~1-1/RC) - 75~. The actual dimension will be smaller
by the cube root or 91% of the ideal compression size.
We say ideal because part of the efficiency loss will
be blow-by or leakage so that the actual volume flow
ratio will be less than 75%. From the expansion engine
86 the now cold air from exit 87 goes directly to the
space being cooled, e.g., an automobile interior, as
cold air~
42~
-27-
Slee~e Valve Piston Air Compressor
It was pointed out earlier that the reduction of
heat exchange between gas and walls was not usually
very important for a standard air compressor because
the heat of CGmpresSion was usually rejected before
use, even though this is inefficient. Thîs assumes
that the gas remains cooler at all times during com-
pression than it would be for the case for a purely
adiabatic compression. On the other hand as explained
above and shown in figure 2, if the compressor cylinder
walls and head are not cooled adequately, then the
gas during compression may be hotter on the average
than would be the adiabatic case and the work required
for compressing a given volume of gas will be greater.
If the cylinder walls and head are sufficiently
cooled, it is possible to lower curve (2) to curve (5)
of figure 2 sufficiently such that T and P are always
less than the adiabatic case (1). This is similar to
what happens with an intercooler between stages of an
industrial air compressor. This requires additional
compressor machinery and the same argument applies to
each sta~e. Therefore, in general it is worthwhile
to reduce heat exchange betw~en the gas and the walls
of any compressor unless the cylinders are especially
effectively cooled. In figure 2 this corresponds to
the isothermal compression of curve 5 that ends up at
the original temperature To.
It is a purpose of this invention to reduce the
heat exchange with the walls of piston compressors by
inducing the inlet gas charge to enter in a near-
laminax flow condition.
To achieve near-laminar flow in a piston compres-
sor the area of the inlet ports is made nearly equal
to ~he area of the piston and as symmetrical as pos-
sible with respect to the axis of the cylinder. Onedoes not wish to induce eddies that circulate the gas
12~)'1290
--28-
rapidly from the walls to the inside volume and back
to the walls ayain, etc.
Near-laminar flow in this context means that
nowhere are the gas flow velocities within the cylinder
significantly greater than the piston velocity. There-
fore within one stroke the gas will not move very
much further in contact with a wail than roughly a
stroke length. If the walls are smooth, this means
that the fractional heat exchange will be small. On
the other hand the gases that leave throu~h an exit
valve can be at a considerably higher velocity than
the piston and exchange heat with the exil: "plumbing",
provided this plumbing does not conduct through a
thermally conducting metal path with too much heat
transferred by conduction to the rest of the cylin-
der. The exhaust stream is all at the same constant
temperature and so the exhaust plumbing can come to
equilibrium. This means that the gases must enter
the cylinder volume with near-laminar flow -- i.e.
slow velocity~ but they can leave more rapidly and
turbulently. Therefore the suction ports must be
large -- at least 1/2 the area of the piston -- while
the exhaust exit port can be smaller.
In an air compressor that uses air operated reed
or spring valves or equivalent for SUCtiOh, it is
almost impossible to obtain laminar flow because the
air operated valve remains open only when there is
continuously a significant fraction of an atmosphere
pressure drop across the valve to overcome spring
tension and inertia. Consequently when the inlet gas
gets beyond the restriction of valve lip, it will
expand and move into the cylinder volume at some sig-
nificant fraction of sound speed -- say 1/2 to 1/4 of
Cs. This in general is 20 to 100 times the maximum
velocity of the piston and so ensures high turbulent
heat transfer during the stroke. The way to avoid
Z~
-29-
this high heat transfer is to induce the gas to enter
the cylinder at the periphery of the cylinder walls
through a large port area, as is shown in figures 8a
and 8b for the case of an air compressor with an inlet
at top of strokeO The area of the port at the cylinder
periphery is: port area = 2(pi) R L, where R = radius
of cylinder, and L = length of port. The piston area =
(pi)R2~ Therefore the port length necessary such
that port area = (piston area) becomes L = R/2. Since
the stroke is 2 R , i.e. stroke = diameter, then the
inlet port length can be a small fraction (1/4) of
the stroke length and the piston rings need not overlap
the port. The flow pattern for a straight-in, i.e.
radially oriented flow~ is shown in figure 9a.
In the embodiment shown in figures 6 and 7b a
piston 91 rides inside a cylinder 92 with a standard
head 93 and exhaust valve 94 that may be any of the
standard types (reed, spring loaded, flapper, etc.).
A sliding ring or sleeve inlet valve 95 opens and
closes an inlet port 96 opening 360 around the cylin-
der and of a height of roughly half the radius of the
piston.
One means of operating the sleeve valve 95 in
proper phase with the piston is to have the sleeve
valve 95 enter a small recess 99 in the head 93 for
sealing the compressed gas from leakage back into the
plenum 100. The plenum 100 carries the induction gas
or air to the inlet valve 95. Vanes 101 at the
entrance to the plenum 100 direct the induction gas
to the inlet port 96. The valve is opened and closed
by a cam 102, and a rocker arm 103 causes a small
rotation of the sleeve valve 95. Ricardo (1954) has
shown the effective opera~ion of sleeve valves in
gasoline and diesel engines. Thousands of British
aircraft engines were manufactured in World War II
with sleeve valves. These valves opened both suction
-30-
and exhaust passages by rotation as well as axial
motion, and the mechanical technology of sleeve valves
for piston machines exists. The opposed porting of
induction and exhaust did not, however, allow near-
laminar flow and introduced about as much turbulenceas the overhead valve, but valve operation was highly
reliable. The small rotation at top and bottom of
the stroke reduced friction by eliminating stiction.
If the vanes 101 that guide the entrance of the
gas into the plenum 100 and the cylinder 92 were radi-
ally oriented as shown in figure 7A, then inflowing
gas will tend to form an annular vortex just like a
large smoke ring (see figure 8A). If the inlet port
area is equal to the piston area, as recommended,
then the velocity of this vortex ring will be roughly
that of the piston.
Simplistically we might think that the vortex
would make roughly 1 turn during each half stroke or
2 full turns during both induction and compression.
Since gas at high Reynold's number must move roughly
50 to 100 diameters to exchange its heat with a wall,
we might believe that 4 turns might be adequately
few. Unfortunately it is not this simple. A vortex
is like a weight on the end of a string that con-
tracts or a figure skater contracting his arms in a
spin. As the vortex is compressed in the compression
stroke, it will spin faster, provided the friction
with the wall is small enough. The friction with the
wall must be made small if the heat transfer is to be
small; the two go together. In the case of the radial
or smoke ring vortex, the velocity will increase as
it is compressed. Conservation of angular momenium
as the vortex is compressed one-dimentionally would
increase its velocity as:
vvOrtex vO(SO/Smin)
where Smin is the compressed stroke length and SO the
maximum stroke length. Hence if the compression ratio,
Z~3~
-31-
SO/Smin, is large, the velocity of the vortex will
increase significantly. Actually if the compression
ratio is large such that Smin less than R, then the
single vortex will break up into smaller vortices as
shown in figures 8B and 8C. Figure 8A shows the single
large annular vortex at bottom stroke; in 8B it is
partially compressed; and in 8C it is fully compressed.
The break up of the single vortex of figure 8A is
shown as four vortex rings in figure 8B and as eight
vortex rings in figure 8C. Let us suppose the compres-
sion ratio is 4:1 as it would be for an industrial
air compressor supplying 100 PSI air. Then the vor~ex
velocity would be increased by two times at maximum
compression and the size (diameter) of each small
vortex would be r/4 and so each one would make a revo-
lution in 1/8 of a stroke period. The result would
be nearly that of turbulence because the vortices
would have time (number of turns) to break up. As a
consequence a significant heat flow from vortices to
wall (particularly the piston and cylinder heads)
will take place. It should be further pointed out
that if the inlet gas had been made turbulent from
the start, i.e. the single large vortex had been made
a larger number of smaller random votices, then the
turbulence as such will be compressed and increases
in strength just as a 3-dimensional gas. (The radial
vortex acts as a 2-dimensional gas.~ Hence the tur-
bulent velocity would increase as
(Volume) 1/3 or: energy is proportional to
(Volume)~2/3
just as would be the case for a gas with a G value
(ratio of specific heats) of G = S/3. The heat
exchange with the walls of the cylinder and heads is
greater because of this turbulent energy increases
with compression.
~34~
-32-
It is an object of this invention to reduce both
the turbulent as well as the single large radial vortex
heat exchange in piston compressors and engines by
means of both laminar flow and the induction of a
weak axial vortex.
This discussion of the formation of an annular
vortex, its compression, intensification, break-up,
and dissipation has been theoretical and somewhat
speculative, but there is now ample evidence from
measurements using doppler laser tracking and finite
element modeling to substantiate this description~
The numerical modeling of Gosman, Johns, and Watkins
(1978) shows all four of the above sequences with
special emphasis on the scale and eddy viscosity.
The finite element size truncates the calculation at
a finite scale larger than the expected laminar bound-
ary layer, but the resulting eddy viscosity gives
vortex life-times in agreement with observation. The
intensification of the primary annular vortex with
compression and its break-up to isotropic turbulence
is also predicted. The fact that the inlet flow
dynamics is the entire source of the turbulence is
also substantiated. It is particularly encouraging
to observe how the experimental measurements of the
in situ cylinder flow so perfectly matches the finite
element calculations. Morse, Whitelaw, Yianneskis
(1979) used doppler laser anemometry to map the flow
patterns in motored piston-cylinder assemblies. These
observed flow patterns of Morse et al. (1979, p. 215)
are the ~omplete confirmation of the theoretical ones
of Gosman et al. (1978, p. 102). I therefore feel
confident in predicting these flows analytically and,
furthermore, defining the way to stabilize them. This
is the axial vortex with near-laminar flow.
f;~
--33~
An Axial Vortex in Piston Machines
If, as shown in figure 7B, the vanes 101 of the
plenum 100 leading to the inlet port are given an
angle with respect to the radial of roughly 60 to
A5, then the entering gas or air will be given a
velocity component tangential to the cylinder walls
as well as radial and an axial vortex will be estab
lished. Figure 9A shows a side view during inlet at
half stroke and the rotational path given to the gas
in the cylinder 2. Figure 9B shows the piston at
bottom of stroke and the rotational gas path. The
axial vortex does not change velocity when it is com-
pressed or expanded axially because its angular momen-
tum is not changed. However when the gas that is
injected at the radius is forced towards the axis by
subsequent injected gas following, the classical vortex
relations hold for a Rankine vortex and the vortex
"spins up", i.e. rotates faster near the center than
at the periphery. Conservation of angular momentum
requires that the tangential velocity increases as
Vtangential = Vo(Ro/R) where VO is the tangential
velocity at the outer cylinder wall of radius Ro and
R is some smaller radius. If we use our standard
port area equal to the piston area and a vane angle
of 45, then the average tangential velocity = 1.5
Vmax where Vmax is the maximum piston velocity. At a
radius R - Ro/2 the rotational velocity is twice
greater than the periphery velocity. This increase
in velocity as the gas approaches the axis is a kind
of centrifugal barrier. It is well known in the geo-
stophic flows of the atmosphere, and is why vortices
are recognized as such stable structures. The point
is that a relatively weak axial vortex will prevent
the annular smoke ring vortex from forming. The cen-
trifugal barrier inhibits flows that interchange fluidelements in a radial direction. As the gas loses
2~ Z 9
-34-
some angular momentum by some friction with the wall,
it finds itself subject to less centripetal barrier
(i.e. less angular momentum) and can therefore more
easily approach the axis. However the radial motion
towards the axis that rorms the smolce ring vortex of
figure 8A is prevented. Instead one has a higher
rotation rate and velocities near the axis. There
will of course be additional friction at the heads
where the azimuthal velocity is greatest, but here
the area is much smaller and so the total heat exchange
is less. For example, at half radius, R Ro/2, the
tangential velocity is 3 Vmax, yet the area is 1/4 of
the head area which in turn is 1/12 of the total area
of cylinder walls and heads for a stroke ]ength equal
lS to a diameter or SO = 2 Ro~ The region of high veloc-
ity of the axial vortex, then, makes only a small
contact with the walls, and therefore there is a very
much smaller heat loss than would occur either with
high velocity isotropic homogenous turbulence, or
with the induction of a large smoke ring vortex. It
is therefore recommended that the laminar flow inlet
valve incorporate a pitch to the vanes that gives the
entering gas a rotational motion such as to form an
axial vortex in order to reduce the heat loss both to
the cylinder walls as well as to the head.
De~ailed Desi~n of Induction Axial Vortex Flow
We have shown that an annular or radial flow
vortex in general adds to the convective transport of
heat and therefore should be avoided. The primary
purpose of the axial vortex is to suppress radial
motion in the gas flow. Hence we desire that the gas
entering the cylinder has as small a radial flow as
possible, so we make the suction port opening as large
as practicable to keep the absolute veloci~y small,
regardless of whether it is radial or tangential, and
2~q3~
--35-
in addition we desire that the flow through the suc-
tion port be at constant velocity both tanyential and
radial. If this were not the case -- for instance,
if we kept the radial flow velocity constant and
allowed the tangential velocity to decrease -- then
the resulting axial gradient within the cylinder of
angular velocity would induce an axial circulation
and induce an annular vortex which is not desired.
The radial velocity is determined by (piston
1~ velocity) x (piston area/port area). The piston area
is fixed. Then velocity = frequency x R sine: (The
projection of the crank arm slightly modifies the
purely sine behavior); we want (piston velocity/port
area) = constant, or port area proportional to sine.
This establishes the cam design that opens and closes
the annular suction port, i.e. the suction port opening
displacement should be approximately (DR) sine, (DR =
maximum suction portion opening). In order that the
annular velocity be held constant during induction we
have several choices.
1. The vanes that give the radial motion to the
induction air can be arranged to give constant tangen-
tial velocity despite the changing port opening.
2. The induction air can be drawn from a plenum
that has a secondary vortex within the plenum that
lasts long enough so that the angular momentum of the
induction gas remains constant during each induction
period.
The second alternative is probably easier because
3~ the secondary vortex is easy to establish and has a
long decay time relative to the piston cycle.
For an example of a conservative design: Let
the tangential velocity be (pi)/2 times the radial
induction velocity and let the radial induction veloc-
ity be the piston velocity. We assume a stroke oftwice the radius. Then:
-36-
S - stroke = 2. R
Vp = piston velocity at mid-stroke, i.e. maximum
velocity (neglecting the crank arm
projection)
ts = time of stroke = (2R/Vp)(pi)/2 = (pi)R/Vp
tcy = time of cycle (2 cycle) = 2(pi)R/Vp
VR = radial velocity of induction air assumed = Vp
VT = tangential velocity of induction =
(pi)VR/2 = ~pi)vp/2.
The time for induction air to make one revoluti.on
= tT = 2(pi)R/VT = 4R/Vp.
The number of revnlutions of air during a cycle
= tcy/tT ~ [2(pi)R/Vp] [4R/Vp] = (pi)/2.
Therefore the expected damping or drag on the
vortex as previously discussed should be fd = #
revolutions/50 - 3~. The fractional heat transfer
should be roughly the same.
Let the plenum surrounding the suction port be
R long and in the axial direction and 2R ln radius.
Then the tangential velocity:
Vplenum at the plenum radius 2Rp becomes
plenum = 1/2 VT = ~(pi)/4]V
and the time to make a revolution becomes
t 1 = 2(pi) Rplenum/Vplenum P
and the number of revolutions in the plenum per
stroke =
tCy/tplenum (pi ) /8 .
This means the vortex will not decay significantly
stroke to stroke since the vortex decay time with the
vanes might be 5 to 10 revolutions.
The plenum will replenish a cylinder volume of
S(pi)R2 = 2(pi)R3 per cycle
and itself contains a volume
R[4~pi) R2] = 4(pi)R3.
-37-
Therefore half the pLenum volume is replaced per
cycle, and the vortex in the plenum will make (pi)/4
revolutions per filling time. This is also adequately
small to ensure constant angular momentum or negligible
decay of the plenum vortex. The input to the plenum
vortex could be a single or several tan~ential ports
of area sufficient to match the mean induction veloc-
ity. The plenum vortex averages the induction veloc-
ity. The mean induction velocity, i.e. average piston
velocity, - 2vp/(pi). The plenum tangential vortex
velocity will be half of the cylinder induction tangen-
tial velocity, or
plenum = VT/2 = ~(pi)/4]V .
The plenum inlet port area is determined by the
relation:
(plenum port area) x (plenum tangential
velocity) =
(piston area) x ~mean piston velocity)
or plenum port area = (pi)R Vp/Vplenu~ = ~/(pi)]R .
Since the plenum is R in cross section, this means
that several ports would be required. A possibly
reasonable design would be four tangential ports of
R/(pi) in width and R long.
An experiment has been performed in which a com-
mercial air compressor was fitted with a clear plasticsleeve valve and clear plastic head so that the flow
could be visualized with smoke. When the plenum inlet
vanes were given the angle prescribed by the analysis
given above, then movies of ~he gas motion show the
expected vortexr If the vanes were oriented radially
no axial vortex was formed and a greater degree of
random motion -- turbulence - was evident.
OPeration of the Sl_eve Valve Piston Compressor
On the intake s~roke air is drawn into the cylin-
der through the oblique peripheral vanes 101 causing
~29~
-3~-
an axial circumferential flow vortex to be formed in
the plenum 100 (see figures 6 and 7~). When air from
the vortex is drawn into the cylinder through the
induction port 96, it forms a near-laminar axial vortex
in the cylinder. The near-laminar flow and stability
of this vortex flow reduces the heat transfer from
the air to the cylinder walls during the rest of the
cycle. The induction sleeve valve 95 begins to open
just after top dead center and is fully open at half
down stroke. It is during this down (induction) stroke
that the near-laminar induction vortex is created in
the cylinder 122. As the piston 91 approaches the
bottom of the stroke, the sleeve valve 95 closes,
thereby maintaining a constant induction velocity as
the piston slows down.
At the bottom of the stroke the sleeve valve is
closed, and compression starts as the piston 91 starts
upward in the cylinder. Compression of the ~as con-
tinues until the gas pressure exceeds the delivery
pressure and the exhaust valve 94 opens until the
piston reaches top dead center. Meanwhile the sleeve
valve is starting down. The induction port 128 opens
a few degrees after top dead center when the small
residual gas in the top-of-stroke clearance has
re-expanded to the induction pressure. During the
further downward stroke, the induction gas is drawn
through the ~leeve valve induction port from the
plenum with its vortex motion. This is now the start
of a new cycle. The reduced turbulence of the near-
laminar flow and the stability of the axial vortexcombine to create a greatly reduced heat transfer to
the cylinder walls and head and piston crown.
Internal Combustion Engine
Typical classical text books on internal combus-
tion engines, like "The Internal Combustion Engine in
., - .~ ~
-.
,
~.`
.,
-39-
Theory and Practice" by Charles F. Taylor (19fi63 or
"The Highspeed Internal Combustion Engine" by Sir Harry
R. Ricardo (1953), make no mention of the transient
effect skin depth heat exchange discussed earlier in
conjunction with thermal skin depth concepts. Instead
running averages are taken for pertinent quantities
and this major inefficiency factor is neg]ected. The
observation that the gas temparature at the end of
compression is close to the expected adiabatic tem-
perature is sufficient ko dismiss the topic, but asdemonstrated in figure 2 and the ensuing discussion
the heat or energy loss in compression can be severe
yet still have the final pressure or temperature some-
times not much different from the ideal adiabatic
case. As a consequence the origin of heat loss from
the gas to the walls is treated in the turbulent flow
limit. The presumption of turbulence is ad hoc. A
necessary condition for turbulence is a large Reynold's
number and indeed this condition is satisfied. How-
ever, the condition for turbulence to uniformly fillthe cross section of the cylinder volume is not
treated. If the inlet valve area is constructed such
that the gas rushes in at a velocity great compared
to the piston velocity, then indeed there is plenty
of time for the high velocity gas to create isotropic
near uniform turbulence within the cylinder volume,
and indeed this is what usually happens. We have
shown instead that the gas can be inducted in a smooth
near-laminar fashion and can therefore grea~ly reduce
the turbulence. Furthermore we can give the gas rota-
tion about the axis, suppressing undesirable heat
exchange vortices. On the other hand many textbooks
and recent designs of internal combustion engines
emphasize the advantage for more complete combustion
by the induction of greater turbulence, especially by
"squishn. The notion is that the turbulence enhances
~X~ L,Z9~
-~o-
the mixing of burning and unburnt gases, thereby pro-
moting greater combustion. This is in conflict with
our objective of decreasing the turbulence in order
to reduce heat transfer.
There are two kinds of positive displacement
internal combustion engines. The Otto cycle involves
mixing fuel with the air before induction, and so
fuel near the cold walls re~uires turbulent mixing to
carry it to the hotter, interior, burning region.
The second type of internal combustion engine is the
diesel, in which the fuel is injected as liquid drop-
lets after the air is compressed to a high temperature.
In this case the axial vortex has a peculiar advantage.
Solid or liquid particles of higher density tend
to centrifuge outward in a gas vortex. The injection
velocities of several-micron-size particles, typical
of fuel injectors, after traveling a centimeter and
slowing down, are comparable to the vortex velocities
expected (several x 103 cm second 1) at radii of Ro/2
to Ro/4. Therefore if the diesel fuel injector is
centered in the head and injects a typical radial fan
of droplets, the axial vortex has the ideal property
of transporting the drops radially outward until they
vaporize and burn. By adjusting the droplet size
distribution and injection angle, the droplets will
automatically be transferred to regions where they
will be most effectively burned. If they have not
been burned in one radial zone, they they will tend
to be centrifuged outward to regions of more air and
less fuel which augments combustion. A further bene-
ficial effect occurs by injecting the fuel near the
axis of a vortex. The fuel burned near the axis of
the vortex will be the hottest region because of the
most complete use of oxygen. However, it is just
this region that is shielded from the walls by the
outer layers of the vortex. Finally it is the nature
-41-
of a vortex to be stabilized by heat such that the
hot region of flow remains stable on the inside. (The
hot air is lighter and "floats" to the axis. In the
centrifugal force field of a rotating vortex the axis
is the gravitational "high" point.) This effect
greatly reduces the wall area in contact with the
hottest gases and consequently reduces the heat flow
from the combustion gases to the walls. Therefore it
is an object of this invention to greatly reduce the
heat flow between the gases and the walls both during
compression and combustion for both the Ot:to and diesel
cycle engines. In the diesel cycle engine it is a
further objective to maintain combustion partially
removed from the cylinder walls thereby further reduc-
ing the heat loss and emissions.
Heat Loss in a ~arnot Cycle
The theoretical maximum efficiency of a compres-
sion engine from the second law of thermodynamics is
just:
Eff = (T - T )/T
where Tl is the initial temperature of the gas and T2
i5 the temperature of the gas after an ideal adiabatic
compression. Since the temperature ratio
T2/Tl = RC(l G)
the ideal efficiency is just
1 - R (1-~)
where R~ is the compression ratio and G is the ratio
of specific heats~ ~ = 1.4. Thus the typical Otto
cycle engine with 8:1 compression ratio should give
56~ efficiency and a diesel with twice the compression
ratio of 16:1 should give 67% efficiency. In practice
the efficiency may be 1/2 of these values. It is
well recognized (Taylor, 1968) that heat loss to the
walls is a major loss, about 30%, and the remaining
loss is ascribed to (1) "time" loss (15%~ which is
-42-
the delay in burning of the fuel relative to the time
of maximum compression and (2) friction of the moving
parts (5%). The skin depth loss discussed in this
disclosure ;s not mentionecl in enyine literature.
The time dependent heating of the walls when averaged
indeed gives the heat loss that is measured in the
cooling water (or air). However, the skin depth
phenomenon also gives rise to an increase in the tem-
perature of the exhaust gases over and above what
they would be ~or an ideal cycle. Let us illustrate
how that can happen.
Exhaust Heat Loss
In general experimental engine measurements do
not compare exhaust gas temperature to the expected
value because of the unknown heat exchange in the
exhaust gas ducts. We consider the efect on exhaust
gas temperature of an ideal pessimistic skin depth
loss cycle. Ambient air enters at temperature Tl and
is adiabatically compressed to T2. Heat is added by
burning fuel to reach a temperature T3 and the hot
gas is adiabatically expanded to T4. The pressure,
volume, compression ratio, temperature and specific
heat relationships are expressed thus:
Voll/Vol2 = R
T2/Tl = RC(l ~)
P2/Pl = Rc
The heat added by burning fuel, H, leads to a
temperature T3 such that
3 2
and the peak pressure is:
P3 = P2 (T3/T2).
The expansion ratio is usually the compression
ratio Rc (without an exhaust turbine) and so the
exhaust gas temperature T4 is related to the peak
temperature T3 as:
T3/T4 = T2/Tl = RC(G 1).
2~3~
-43-
Skin Depth Loss
Let us suppose a fraction of 40% of the heat at
T3 is stored in the cylindrical piston wall at the
time of burn. Then T3 = 0.6 T3 and the new values of
P and T and P' and T', become:
P~' = (0.6 T3/T2) P~
and the exhaust temperature becomes
~4 = 0.6 T3 ~c(1-5)
Let us suppose that of the 40~ heat loss to the wall,
1/2 is returned to the inlet gas during induction an~
1/2 is lost to the cooling water. In other words,
the thermal skin depth thermal wave divided equally
between in-going and out-going heat flow. We recognize
that some of the heating of the inlet gas will take
place after compression starts and give rise to the
mechanical loss discussed previously but we suppose
for simplicity that this is smaller than the initial
heating before closure of the inlet valve and start
of the compression stroke. Then the compression will
start with the conditions
Tl' = Tl + 0.2T3
Pl = Pl
and the conditions after compression become
- (Tl + 0.2 T3) R (G 1).
The same amount of fuel is burned as before so that
T3' = T2 + (T3 T2)
and
3 P2 (T3 /T2').
Again a 40% loss takes place which is now more heat
into the wall since T3' is greater than T3. In each
succeeding cycle T3'' is greater than T3' is greater
than T3 etc. This sequence of increasing temperature
must be limited by secondary effects.
Let us calculate typical values for a diesel
engine with Rc = 16; T3 = 2~2; so that the heat added
-44-
is H = T2~ This corresponds to a ~uel mixture corre-
sponding to 26% stochiametric. Then:
T = T R (G-l) 3 T
Tl' + 02 T3 = Tl ~ 0.4 T2 2
so that
T2l = 2.2 T2
and
3 T2 + T2 = 3- 2 T2 = 1. 6 T3.
The exhaust temperature then becomes:
T4' = 0.5 T3' RC(l ) = 0.96 T4.
This is the expected value of the exhaust temperature
without the 40% heat loss, but less work is performed.
In other words the temperatures increase by a ratio
of 1.6 per cycle and the peak pressure
P~' = (1.6~2.2) P3 = ~.73 p3
decreases 0.73 per cycle. Hence more heat is lost
and less useful work is done each cycle until another
limiting effect takes place. One such effect is that
T2, the compressed input gas temperature, becomes
greater than 0.6 T3, i.e. the wall-cooled combustion
gas temperature. In this case th~ heat loss during
the compression stroke will limit the subsequent com-
bustion gas heat loss because the wall will be already
preheated. If T3 then remains the same, let us say
25 lo 6 T3, and the heat loss to the skin depth changes
to 20~ of T3 (another 20% is loss from T2), then the
exhaust gas T4 will increase to 1.2 T4 and 20~ of the
possible useful energy will be lost in the exhaust
rather than through cooling water. Natura:Lly these
relationships are extremely complicated and a very
detailed computational analysis needs to be made to
give accurate predictions/ but the above analysis is
enough to give the direction of how to correct for
these losses. The direction of these losses helps
explain why the diesel engine can not be made more
efficient by increasing the compression ratio further,
9~
-45-
say from 16 to 20. It i5 observed (Taylor, 1966)
that the useful work or efficiency remains constant
as a function of compression ratio in the range greater
than 16. The reason for this is that as the compres
sion ratio increases, the heat loss increases because
the geometry of the clearance volume at the end of
the stroke becomes thinner so that the sur~ace to
volume ratio becomes larger and so heat losses are
larger.
Diesel En~ine Design
We propose a geometry for a diesel engine that
greatly reduces:
1) thermal loss to the walls.
23 time loss of compression ratio during
combustion.
3~ mass of moving parts.
The geometry combines the concepts of near-laminar
flow and separate supercharging, combustion, and expan-
sion cylinders (2-cycle~.
The overall compression ratio we choose is 20:1.
The first supercharging cylinder has a volume ratio
of 5:1~ Therefore the combustion cylinder has a volume
ratio of 4:1 so that the net compression ratio is
20:1. The first expansion in the combustion cylinder
has the same volume ratio of 4:1. If the combustion
leads to doubling of temperature and hence pressure,
then the expansion cylinder must have a volume ratio
of
5 x 21~G = ~.2 1
so that the exhaust pressure is reduced to the ambient
atmospheric pressure.
The advantages are:
1. The supercharging cylinder and piston are
subject to a lower pressure than the peak combustion
pressure by the ratio of 1/(2 x 45) = 1/14 = 7%. It
29~
-46-
can be made correspondingly lighter, with a shorter
piston skirt, short stroke and larger diameter.
2. The combustion cylinder for equa:L length
stroke will be smaller in diameter by 5-1/2 = 1/2.24 =
45%, or the piston area will be smaller by 1/5. Hence
the maximum force on the piston and head will be
smaller by the same ratio, i.e~ 1/5, than a cylinder
of a standard diesel of the same power and stroke.
Therefore the mass and friction will be smaller by
the same ratio.
3. The compression ratio of the combustion cylin-
der is only 4:1 and we assume a long stroke of 4r or
two diameters. If the bottom port length is 0.5r
(2-stroke), then the compression or expanCion stroke
is 3.5 radii long. This leads to a large crank angle
of peak compression. For example, we consider combus-
tion taking place over the range of compression ratios
16:1 to 20:1 to 16:1l or 67% to 70~ to 67~ ideal effi-
ciency. Then this translates into a compression ratio
in the combustion cylinder of 3.2 to 4 to 3.2. The
total crank angle corresponding to these displacements
is 54 degrees or about 1/6 of cycle. This is larger
than the equivalent single cylinder 2-stroke diesel
where for Rc = 16:1 to 20:1 to 16:1 corresponds to 24
degrees or a factor of 2.25 shorter time~ Therefore
the combustion time compression loss is correspondingly
reduced.
4. The combustion piston-cylinder geometry during
peak compression and combustion period (16:1 to 20:1
to 16:1) is roughly 1 radius long. This is the mean
head clearance during combustion for a 3.5 compression
ratio for a stroke 3.5 radii long. Therefore the
surface to volume ratio is favorable to reduce heat
loss to the walls by a factor of (1 + r/z) = 5 fold
compared to the standard single cylinder mean head
clearance of 1/9 radius Eor a 20:] compression ratio.
12~?4~
-47-
5. The expansion cylinder can be made larger
than the supercharging cyl;nder so that the exhaust
gases can be "over-expanded", actually correctly
expanded, to extract all the useful work of the
exhaust. So-called over-expansion of exhaust means
that the exhaust gas is expanded down to atmospheric
induction pressure. Normally the exhaust pressure is
2 to 2.5 times induction and the gas energy is either
used inefficiently, 50% efficient in an exhaust turbine
and superchar~er, or wasted entirely. The separate
exhaust cylinder allows for the correct over-expansion
as well as the advantages of the smaller size and
smaller compression ratio of the combustion cylinder.
6. Finally by careful port design we can reduce
the usual large turbulent heat exchange with the walls
by establishing the near-laminar flow axial vortex in
each cylinder. This will require that the transfer
of the gases from each volume be quasi-static and
that, therefore, the pressure drop across valves be
very small.
Separate Cylinders for
SuPercha_~-n~_Combustion, Exhaust
The primary reason for the three separate cylinder
designs is to limit the compression ratio of the com-
bustion cylinder to a small enough value such thatthe cylinder piston geometry during combustion is
roughly a "right" cylinder, i.e. where the length
equals the radius so that the near-laminar flow axial
vortex can effectively reduce heat flow to the walls.
The reduction of combustion time compression loss and
reduction in mass are additional benefits. On the
other hand two additional cylinders and complex valving
must be added. The major useful work is recovered in
the exhaust cylinder. This mechanical energy is larger
than the combustion cylinder by the temperature ratios
4;~9~3
-48-
(8.2~(G 1) to ~3.5)(G 1) or 2.32/1.65 = 1.41 fold-
Therefore the heat losses in the combustion cylinder
are somewhat reduced in importance compared to the
exhaust cylinder design. As a consequence we design
the combustion cylinder with some compromise of lami-
nar flow conditions.
Combustion Cylinder Design
The ideal laminar flow compression or expansion
cylinder has been described above. The induction
port is at the top of the cylinder wall, is r/2 long
and allows slow laminar azimuthal and radial flow.
This slidin~ wall valve would be difficult to seal
and cool at the extreme temperatures, 1500C, and
pressures, 1800 PSI, of combustion. However, Ricardo
(1953) has shown that a two-stroke slide valve diesel
can be made to work well, but in the present case the
induction air is much hotter (300C)~ In addition
scavenging is strongly favored by an axia:L flow through
the cylinder from bottom to top. We therefore pro-
pose that the combustion cylinder be made similar toa 2-stroke diesel where induction takes place with
the supercharged gas (5:1 compression, 10:1 pressure)
in the usual annular ports at the bottom of the stroke.
These ports will be r/2 long, the same as the sleeve
port allowing induction for 83 degrees. Exhaust leaves
by an axially centered valve in the head.
Exhaust Cylinder Design
Since the exhaust gas pressure leaving the combus-
tion cylinder is roughly twice the pressure of the
induction air, part of the bottom stroke period must
be used for letting down this higher pressure to the
value of the supercharged inlet pressure. On the
other hand the inlet gas must be transferred to the
combustion cylinder and scaveng~ out the exhaust gases.
-49-
We propose to accomplish this with near-laminar flow
by arranging the inlet port and the exhaust port to
be open at the same time. Therefore the pressures
will all be equalized.
Since the volume is changing during this time as
the exhaust piston descends, the induction air and
exhaust gases will be expanding during transfer. The
pressure starts at the exhaust pressure of the combus-
tion stroke. Therefore the induction air must be
over-compressed, above the mean supercharge value,
when the induction port opens. The combined volume
of gas of the combustion cylinder and supercharged
storage volume will expand adiabatically as the ex-
haust piston descends and the exhaust gas will be
replaced by a fresh charge of induction air that
expands back down to the design supercharge value
during induction.
We describe two methods for this let-down. The
first will not work because the time is too short.
Since this let-down must be an adiabatic expansion in
the exhaust cylinder, the crank angle for the exhaust
cylinder for a 2:1 pressure expansion of the exhaust
becomes the let-down time. This let-down time is the
time for the exhaust cylinder to move from top of
stroke to a fractional volume = (1/8.2)(1/2) = 0.0743.
For an exhaust piston stroke length the same as the
combustion cylinder, this corresponds to a crank angle
of arc cos [1-2(.0743)] = 32 degrees. The exhaust
cylinder next must accept an equal volume of scavenge
exhaust gases at constant pressure, i.e. the super-
charge induction air pressure during an angle of arc
cos [1-4(.0743)~ = 45 degrees. The difference in
timing between these two angles is 13 degrees. This
is the time for the supercharged air to displace the
exhaust out of the combustion cylinder. This time is
too short and would require that the supercharging
4;~
-50-
cylinder deliver all its charge at constant pressure
in less than the required displacement time.
Supercharge Holdin~_Volume
Instead, to avoid this problem, the supercharger
can deliver its air at twice induction pressure to a
holding volume. Then during exhaust let-clown, the
pressure drop of two-fold occurs in the combined vol-
ume of supercharge holding volume, combustion cylinder,
and the initial stroke of the exhaust cylinder. At
the same time a full charge of supercharged air is
inducted into the combustion cylinder. The super-
charger cylinder recharges the holding volume (adiaba-
tically) back to its original pressure of x 2 induc-
tion pressure. The holding volume then becomes:
[1/(21/Gl)] ~ 1.56 x Volume of combustion
cylinder.
Then induction will take place over a crank angle of
the exhaust piston of
arc Cos[l-2(l-56/Rc-exhaust)]( / ~P
This time conveniently can be placed in the middle of
the combustion cylinder induction port opening of 83
degrees so that the induction ports are almost fully
open (88%) during let-down, induction, and scavenging.
We next must consider the thickness for the inlet
ports to the combustion cylinder in order to obtain
low induction air velocity and hence low heat loss.
We assume a fractional web thickness of the induction
ports of 20% so that the effective induction area is:
[1-0.2] 10.80] [(pi)r2] = 0.7 (pi)r2.
The mean density of the induction gas is about
[1-1/2(1-2 l/G)] = 0.8 its final density and the
effective area is almost 87% of the combustion piston
area. The fractional induction time compared to a
half stroke is [2/(pi)][40/180] and so the induction
radial velocity will be:
~ ~R 4;290
-5:L-
[2/~pi)][l/effective area][180/40] =
3.3 x piston velocity. ~e choose the azimuthal veloc-
ity of the axial vortex to be 1.5 times the radial
induction velocities. This ratio is determined by
the angle of the webs acting as vanes. In this case
the gas makes:
(velocity x stroke time)/circumference =
about 4 turns
during compression and a like amount during expansion.
The heat exchange should then be 10 to 15% provided
the fuel is burned in the inner 25% mass fraction of
the axial vortex (for 25% stochiametric burn).
Scavenging
The inlet supercharged induction air is centrifu-
gally "heavy" relative to the combustion products.By this is meant that the induction air is cooler and
has a higher angular momentum than the exhaust gas.
It is cooler because it has not yet undergone combus-
tion and it has a higher angular momentum because it
~0 has not had as long to spin-down in contact with the
walls. Therefore it will tend to enter as a thin
layer in contact with the outer cylinder wall, displac-
ing and forcing the hotter exhaust gases towards a
smaller radius and hence towards the exhaust valve at
~5 half radius. The inner 1/8 mass of core of exhaust
gases (1~4 in volume, 1/2 in density) will tend to be
weakly turbulent because of the momentum of the fuel
spray. This favors interaction with the entering
cooler induction gas and favors it being swept out of
the exhaust valve. If this residue of exhaust gases
proves to be too large, i.e. poor scavenging, then it
is a simple matter to reduce the size of the exhaust
valve to lJ3 radius and only 5% of the exhaust mass
would possihly be left behind. The axial vortex with
a control exhaust valve therefore naturally lends
2~C~
-52-
itself to efficient scavenging. Ricardo (1954) had a
problem with scavenging where he introduced "swirl",
i.e. rotational angular momentum, in the inlet gases.
This was because his exhaust port was also at the
cylinder peripheral wall instead of near the axis.
The inlet gas then forced the exhaust gases towards
the axis away from the peripheral port and poor
scavenging was the result.
Diesel Engine Desi~n
In figure 10 three pistons -- the supercharging
201, the combustion 202, and the exhaust 203 -- are
driven by a crankshaft 204 and cranks 205, 206, 207,
in respective cylinders 208, 209, 210. The super-
charger cylinder 208 intake has a cylinder wall sleeve
valve 211 driven by cams 212 on ~he crankshaft 204.
These cams 212 drive the cylinder wall sleeve valve
211 so that it opens and closes an annular port 213
that allows induction air to be drawn from an annular
plenum 214 with rotational circulation created by a
volute inlet passage with straightening vanes 215.
The rotation causes an induction axial vortex 216
(see Fig. 11) in the supercharging cylinder. The
large induction port area 213 and axial vortex 216
result in small hPat exchange with the cylinder wall
208 so that the supercharger air is compressed adiaba-
tically and leaves the cylinder via the leaf spring
exhaust valve 217 which is arranged semi-circularly
in the head 218 so that the axial vortex gases leave
the cylinder at roughly one half radius and in the
general direction of rotational flow. The super-
charged air is compressed 8.2 fold to approximately
280 PSI and transferred in a duct 219 to chamber 220
that holds a volume of compressed (supercharged) air
1.56 times the volume of the combustion cylinder 209.
The walls of the duct 219 and this chamber are insulated
or ceramic lined 221 to reduce heat loss.
g~
-53-
In operation the pressure of this supercharged
air in cylinder 208 and storage volume 220 is the
same as the pressure in the combustion cylinder 209
just before the exhaust release and is also the pres-
sure in the exhaust cylinder 210 at the same time sothat all three pressures are the same. The combustion
piston 202 is just uncovering the bottom port 222 of
the combustion cylinder 209 and the exhaust valve 223
in the combustion cylinder head 224. The gases are
then transferred ~figure 11) from the storage volume
220 to the compression chamber inlet plenum 225 via
the transfer duct 226 at almost constant pressure and
adiabatically and thus very small turbulence is
induced. When the pressure has been reduced by a
factor of 2 from 280 PSI to 140 PSI by the expansion
of the exhaust piston 203 in the exhaust cylinder
210~ the combustion cylinder is charged with fresh
supercharged air and the exhaust valve 223 closes.
The supercharged air is then trapped in the combus-
tion cylinder 209 by the top rings of combustion
piston 202 covering the inlet port 222. rrhe inlet
port 222 is a 360 opening to the cylinder having
many vanes oriented about 60 to the radii of the
cylinder so that the entering gas will make the pre-
ferred axial vortex. The supercharge gas pressure
ratio is 5:1 which is the ratio of area of the super-
charging cylinder 208 to combustion chamber cylinder
209. The further compression of the supercharged air
in the combustion cylinder 209 is 4:1 before fuel
injection so that the total compression ratio is 20:1.
Fuel is injected into the combustion chamber by a
fuel injector 228 of a standard type driven by a stan-
dard fuel pump (not shown) driven from the engine
crankshaft 204.
When fuel is sprayed as fine droplets 227 (see
figure 12) into the axial vortex 229 at the axis from
-54-
the injector 228, the cone of droplets is confined to
the axial region until ~hey are radially centrifuged
towards the cylinder walls. The flame combustion of
the fuel remains in the region of the center of the
vortex and only unburnt fuel droplets tend to escape
radially to regions of pre-combustion with available
oxygen. Hence the combustion progresses radially
outward until the fuel is burned. The hot combustion
products then remain separate from the cylinder wall
209 and only the piston 202 and h~ad 223 are exposed
to the hot gases over a limited area. It is a further
object of this invention to utilize the near-laminar
flow vortex to restrict the combustion to the central
axial volume of the cylinder as well as to reduce the
heat flow between the working gas and the walls.
The tubular exhaust valve 223 is of a special
cylindrical design shown in one form in figure 10 and
another in figure 12. In figure 10 the valve is driven
by an overhead cam 230 and shaft 231 and rides on the
central cylindrical head portion 232 which is carried
by the shaft 231 and is water-cooled in passages 233
so that the high exhaust gas temperature, 1100C,
does not overheat the valve. In addition the annular
opening at half radius allows the axial gas vortex
227 to exit the chamber with minimum turbulence induced
in the exhaust gas duct 234. The duct is lined with
ceramic coating or other high temperature insulation
to reduce heat conduction. The exhaust gas duct 234
is made short and of small volume, a few percent of
the exhaust cylinder volume 210, so that the only
valve between the combustion and exhaust cylinders is
the exhaust gas valve 223.
The exhaust gases enter the exhaust cylinder 210
from the duct 234 through an induction port 236 at an
angle to the radius of roughly 60 so that the axial
vortex 237 is formed. The exhaust valve 223 closes
42~
-55-
when the exhaust piston 203 is 1/5 of its stroke after
top dead center. This is when the exhaust gas is
expanded in pressure from 280 PSI down to 140 PSI.
The exhaust piston is then traveling at 2~'5 of its
maximum speed so that the induction port area 236 is
designed to be 1/10 the piston area so that the enter-
ing gas is moving at 4 times the maximum piston veloc-
ity. Since the axial length of the ports is R/5, the
necessary length around the azimuth to give a port
area of 1/10 the piston area becomes [(pi)/2] or 1/4
turn.
The exhaust piston 203 then expands the exhaust
gases by a total volume ratio of 8.2:1 which brings
the pressure down to atmospheric pressure at the end
of the stroke. An exhaust valve 238 is of the same
design as the exhaust valve 224 of the combustion
cylinder 209. It is opened by a cam 239 driven by
the cam shaft 231, the same as the combustion cylinder
exhaust port. The exhaust valve 238 stays open until
the exhaust piston 203 is just before top dead center.
The lead angle of the exhaust port is such that the
trapped gases are compressed up to the value 280 PSI
in the combustion cylinder 209 just before the inlet
ports 222 and exhaust port 224 are opened. The exhaust
gases leave via the volute 241 and the exhaust port
242.
As shown in figure 12 the head portion 250 of
the combustion chamber -- likewise the exhaust chamber
-- can be supported within the sleeve valve 252 by
lugs 2S4 that extend out through slots 256 in the
valve, and the valve can be operated by a rocker arm
258 through a tappet head 260 on the valve.
The timing diagrams, figure 14, show the relative
timing of the pistons and valves. Starting with top
dead center of the supercharging piston 20L, the induc-
tion sleeve valve 211 is 90 ahead and just about to
~4,;~9~
-56-
open. The combustion piston 202 is at 138 just
before the combustion exhaust valve 224 opens and the
combustion piston 202 uncovers the induction bottom
ports 222. The exhaust piston 203 is also at top
dead center. At a later time, 82, the bottom combus-
tion ports 222 close and at 180 the supercharging
induction port 213 closes. Compression takes place
in both supercharging and combustion cylinders 208
and 209. At about 20 before the top dead center of
the combustion cylinder 202, fuel injection commences,
starting the combustion processes followed by
expansion.
Otto Cycle Engines
As previously discussed, an Otto cycle engine
(i.e. gasoline and carburetor) usually is designed
with maximum induced turbulence. When the piston
crown or head is designed with a smaller diameter,
re-entrant volume and then the clearance between the
outer radius of the piston and head is made small,
the gases are "squished" into the re-entrant volume.
This is called squish because the gas in the small
clearance volume is "squished" into the re-entrant
volume thereby inducing turbulence at the end of the
stroke. The objective of this turbulence as explained
before is to ensure that the air fuel mixture, in
contact with the walls, is continously mixed with the
hotter burning interior region so that combustion is
more complete and less unburnt products result in
causing pollution. However, the turbulence increases
hea~ loss.
We have already discussed a diesel engine in
which the fuel is introduced at the axis of the
laminar vortex, thereby shielding the fuel from the
walls and resulting in greater burn efficiency. In
the diesel the fuel is injected late just as it is
-57-
required for combustion. In the Otto cycle engine
the fuel and air may be pre-mixed, in which case there
will always be air-fuel mixture in contact: with the
walls, or conversely fuel-injected engineC; allow the
possibility of a "stratified" charge. If the fuel is
injected along the axis of the laminar vortex before
the top of the stroke, then it will mix with air only
out to a prescrihed radius determined by the droplet
size, vortex velocity and evaporation rate. The
resulting stratification should then be enough to
prevent the air-fuel mixture fr~m reaching the outer
walls. In this case there will be little or no air-
fuel mixture in contact with the walls and no tur-
bulence is required to give complete burning~ Some
"swirl" is sometimes induced by the shape of the head
valves to create an axial vortex in Otto fuel injected
engines, but the induction velocity is so high from
restricted area ports and the non-uniformity so great
that the resulting flow, although creating a vortex,
nevertheless thoroughly mixes fuel and air out to the
peripheral walls and turbulence is required to com-
plete combustion. Instead, we suggest fuel înjection
into our near-laminar vortex and expect ~hat very
little fuel will reach the air in contact with the
cold peripheral walls. Hence combustion will proceed
with the fuel-rich inner core of the vortex. The
air-fuel mixture in contact with the head and piston
crown at the center of the vortex is in contact with
surfaces that remain hot since they are not cooled by
oil film sliding. We therefore consider the design
of a laminar axial vortex fuel-injected Otto four~
stroke engine where some heat conduction loss is
expected, but because of the laminar axial vortex the
heat loss is less and the combustion loss and hence
pollutants from the wall are reduced.
2~
-58-
The cycle is a standard 4-stroke, and compression
and combustion occur in the same cylinder as compared
to the above-described diesel design of 3 cylinders.
The lower compression ratio (e.g. 8:1) of the Otto
cycle partially circumvents this requirement.
A more efficient design can be made with a super-
charger because the final compressed volume becomes
more favorable (larger head clearance) ancl proportional
to the supercharginy ratio, but even without super-
charging, the laminar axial vortex with combustionmaintained clear of the peripheral walls will signifi-
cantly reduce the heat loss~
Four Stroke Otto En~ine
In figures 15 and 16, a piston 301 in a cylinder
302 is driven by a crankshaft in a 4-stroke mode by a
crank arm 303 and a wrist pin 304. The piston is
shown at top dead center just as combustion takes
place with clearance 305 between the smooth piston
crown and head 306. The head clearance 305 is R/4 (R
is the piston diameter) in dimension so that for a
typical stroke length equal to the piston diameter or
2R, the clearance 305 of R/4 corresponds to a typical
Ot~o cycle compression ratio of 8:1. The compressed
volume is bounded by the sleeve valve 307 that rides
~5 in clearances in the lower cylinder and clearances in
the head 306. These clearances, as Ricardo (1953)
discusses, automatically adjust themselves so that
the sleeve valve 307 expands by heat until perfect
sliding contact with the outside wall is made which
then extracts the heat and limits the expansion. The
sleeve valve 307 extends beyond the piston 301 into a
get-lost-volume 310 in the cylinder wall. This get-
lost-volume is such that it may be zero and therefore
act as a valve seat, or, if larger, such that trapped
gases are not compressed forming squish. The sleeve
4~
~9 .
valve 307 is spring-loaded open and actuated by two
cams 311 -- it can be activated by rocker arms as in
figure 13. The head portion 312 within the valve is
supported by the camshaft but can be supported as in
figure 13~
At top dead center at the start of the induction
stroke, the sleeve valve is retracted to open the
port area 314 to the plenum 315 for air induction as
an axial vortex ~ormed by the momentum of inlet air
straightened by vanes 316 in volute air intakes 320.
The head 306 contains a fuel injector 317, spark
plug 318, and exhaust valve 319. The fuel injector
is an entirely standard type, but it is mounted axially
so that the fuel spray will be axially symmetric and
result in a vortex stratified charge. The spark plug
318 should have a spark ignition surface that is flush
with the head 306 so as to present as small a pertur-
bation to the rotational flow as possible. Such spark
plugs are typical of aircraft ones. Similarly, the
bottom surface of the exhaust valve 319 must be smooth
so as not to perturb the vortex flow. The exhaust
valve 319 is driven by a cam 321.
Figure 16 shows the piston 301, the head 306 and
the plenums 315 in the form of two volutes. The vanes
316 create an induction vortex 324 and an internal
axial vortex 325. The induction area of the volutes
at the vanes 316 is 1/3 that of the piston so that
the induction vortex will have a velocity about 4
times that of the piston at the smaller radius of the
piston. Hence the radial induction velocity of twice
the piston velocity is still smaller (about 1/2) of
the axial vortex and so the axial vortex will dominate
the flow and prevent an annular or radial vortex from
forming.
The timing diagram is identical to any four-
stroke engine and so is not shown since 4-stroke engine
~Z5~ 9
60-
timing is so well known. The only small difference
is the cam shape for opening and closing of the induc-
tion sleeve valve 307. The port area is clefined by
the degree of openin~ of the sleeve valveO This area
should be proportional to the piston veloc:ity so that
the ratio of induction radial to azimuthal velocity
remains constant during induction.
Therefore the valve opening should be proportional
to the sine (of the crank angle). This is the simplest
cam to make and is just an off-center circle. In
this fashion both the angular momentum as well as the
radial momentum of the axial vortex remain constant
during the induction stroke and hence the meridianal
or radial circulation is minimal. This ninimi~es the
heat loss.
VIII Conclusion
A class of positive displacement machinery is
designed where there is definitive control of the
heat flow from and to the working gas and its bound-
aries. The result is a major improvement in thethermal efficiency of such machinery. The method of
control is negation of turbulent heat transfer by the
creation of a quiescent near-laminar flow and more
especially such near-laminar flow in a stable vortex
that takes advantage of the symmetries of the confine-
ment volume.
The attainment of extremely low heat transfer
requires that at all times the induced turbulent
velocities be small compared to the displacement
velocity and the total displacement of the working
gas in contact with a wall be a distance that is small
compared to approximately 50 channel widths. This is
the distance of fluid displacement in contact with a
smooth wall at large Reynold's number necessary to
induce fully turbulent flow. Proper induction port
1~4290
-61-
design and small displacement ensures near-laminzr
flow. Within this principle are the embodiments --
(1) an articulated vane compressor-expander applicable
to an automobile size heat pump; (2) an adiabatic air
compressor; (3) a diesel engine where the separate
functions of (a) pre-compression, ~b) post-compression,
combustion and pre-expansion, and ~c) post-exhaust
expansion are carried out in three separate cylinders;
and (4) two and four stroke Otto cycle engines, pref-
erably with fuel injection.
The heat efficiency of these designs in somecases may be up to a factor of 2 better than current
practice. This is because up until now, the heat
transfer from gases to walls was not explicitly treated
as being sensitive to the level of induced turbulence.
Older textbooks such as Ricardo (1954) and Taylor
(1966) barely even consider the turbulent flow pattern
in engines. Recent modeling and measurements demon-
strate unequivocably the induced turbulent flow pat-
terns. The relationship between this flow pattern,the heat loss, resulting thermal efficiency an~ inally
the necessary corrective measures is the foundation
of the present invention.
9~
-~2
REFERENCES
(1) Books
American Handbook of Physics, 1963, Prentice
Hall, N.Y., pp. 256 and 257.
Ricardo, H.R., 1954, The Highspeed Internal
Combustion En~ine, Blackie & Son, London, EnglandO
Taylor, C.F., The Internal Combustion Engine In
Theory and Practice, 1966, Vol. 1, 2.
(2) Articles
Gossman, A.D., Johns, J.R., Watkins, A.P., 1973,
"Development of Prediction Methods for In-Cylinder
Processes in Reciprocating Engines," Proc. General
Motors sYmp.~ Detroit, Mich., p. 103. (Figure 16)
Morse, A.P., Whitelaw, J.H., Yianneskis, M.,
June, 1979, "Turbulent Flow Measurements By Laser-
Doppler Anemometry in Motore~ Piston-Cylinder
Assemblies," Journal of Fluid Engineering~ Transactions
of the ASME, Vol. 101, p. 215. (Figure 17).
(3) Patents
Brewer et al. U.S. Pat~ No. 3,343,782, 9/26/67
(Rotor end sealing using sealing "washers" and rela-
tionship with the rotor bearings.)
Ezop U.S. Pat. No. 3,346,176, 10/10/67 (Sealing
between the rotor and the stripper landing and using
Molybdenum Disulfide coating on stripper landings.)
Brewer et al. U.S. Pat. ~o. 3,356,~92, 12/5/67
~Notches on the inside wall of the housing to reduce
sudden pressure changes. Bearing shoe and sealing
shoe. Also specific materials and combinations of
vanes.)
Pasek et al. U.S. Pat. No. 3,370,785, 2/27/68
("Impeller" disc air filter mounted to the pulley.)
Adsit U.S. Pat. No. 3,401,872, 9/17/68 (Molded
in place plastic rotor lining to hold sealing shoe.
Bearing shoe is molded with the lining.)
Q
-63-
Brewer et al. U.S. Pat. No. 3,419,208, 12/31/68
[(1) Spot welded metal rotor liners to hold sealing
and bearing shoes and alignment dowels. ~2) Counter-
weights are curved to allow easy assembly., (3) Vanes
molded to counterweight hubs. (4) Vane molded of
reinforced thermosetting plastic. (5) Various fibers
in vane plastic. (6) Ball hearing molded into housing.
(7) Pulley hub smaller than the ball beari,ng to facil-
itate assembly.]
Rohde U.S. Pat. No. 3,437,264, 4/8/69 [(1) Rotor
having an abradable coating to seal at the ends of
the rotor and the housing in a recess. (2) Stripper
landing and the coating relationship. (3) Coating
contains MoS2.]
Rohde UOS. Pat. No. 3,437,265, 4/8/6g (Wedge
shaped space between bases of the shoes and their
holding strips.)
Stiles et al. U.S. Pat. No. 3,844,696, 10/29/74
(Counter balanced two-vane extended rotor unit; shroud
to reduce noise on the intake port.)
Ziehl U.S. Pat. No. 3,954,357, 5/4/76 (Standard
two-vane unit restraining pilot sleeve in a guide
track for the pin on which the vanes pivot.)
Three patents filed in 1965 on three-vane full
length rotor type units made until about 1969.
Five patents filed in 1966 and all diagram the
counterweighted two-vane, extended rotor units that
are presently in production.
