High swirl very low pollution piston engine employing optimizable vorticity

Method and apparatus for producing intense, consistent swirl and air-fuel-E.G.R.-vorticity charge uniformity without any net penalty in engine power, volumetric efficiency, or pumping work. Method and apparatus permit optimal swirl and turbulence for flame stability and fuel economy throughout the R.P.M.-load phase space of engine operation. Method and apparatus require no substantial changes in combustion chamber shape or basic engine structure. For spark fired engines, the method and apparatus permits operation at air-fuel-E.G.R.-intake manifold vacuum combinations having excellent fuel consumption characteristics combined with very low emissions of CO, HC, and NO. For diesel engines, the method and apparatus permits optimized swirl without the volumetric efficiency and pumping work penalties accepted with present art swirl inducing techniques.

It is the purpose of the present invention to satisfy a number of 
requirements commonly thought to be in conflict using structurally simple, 
mass producible, and reliable hardware. The requirements to be met are: 
a. Build an engine producing ultralow NO.sub.x and other emissions without 
catalytic control of No.sub.x. 
b. Build the engine in such a way that the goals of optimal fuel 
consumption and NO.sub.x control are not in conflict, with fuel 
consumption superior to that of pre-emission control engines. 
c. Build the engine such that very high power outputs, substantially in 
excess of the power outputs of current production engines are possible. 
d. Build the engine so that combustion is always very smooth, so that 
excellent driveability and engine flexibility are obtained under all the 
many conditions under which automobile engines are expected to run 
smoothly. 
e. Build the engine as a modification of current engines, in such a way 
that current production lines and techniques require little modification 
to produce the improved engines. 
f. Build the engine modification so that it requires fewer and simpler 
parts than those required for current production engines, to minimize both 
fabrication and maintenance costs. To achieve these goals together has 
involved difficult conceptual and practical problems, and particularly has 
required a fundamental breakthrough in the application of detailed 
turbulent fluid mechanics to the physical chemistry of combustion in 
engines. 
The following considerations, many of them not obvious and out of harmony 
with conventional doctrine in automotive engineering, were involved in the 
conception and development of the present invention: 
1. The commonly held "trade-off" between low NO.sub.x emissions and optimal 
fuel economy is based on the narrow range of flame stability limits 
characteristic of prior art spark fired engines. Thermodynamically, 
theoretical efficiency of the constant volume fuel-air cycle for a set 
compression ratio continues to improve monotonically as air-fuel ratio is 
leaned from the stoichiometric ratio. The chemical kinetics of NO 
formation in engines is such that the maximum NO concentration is formed 
around 0.9 stoichiometric (0.9 equivalence ratio), but as the mixture is 
leaned beyond 0.9 equivalence ratio, NO decreases; and leaner than an 
equivalence ratio of 0.65 NO.sub.x is negligible even without dilution of 
the charge with residual gases (products of combustion). With exhaust gas 
dilution, NO formation rates become negligible at richer ratios, and again 
the constant volume cycle efficiency under part loads improves as dilution 
of the charge is increased. With current engines there is a trade-off 
between low NO and fuel efficiency because the very lean or dilute 
mixtures required for low NO burn so badly that the thermodynamic 
advantages are overshadowed by the unstable and slow combustion. The 
conventional wisdom of automotive engineering is that the very lean 
mixtures required for low NO must always burn badly, and so the idea of an 
inherent trade-off between NO emission control and engine efficiency has 
become entrenched. The inventor proceeded in engine research for more than 
eight years to achieve adequately fast and very stable combustion of the 
very lean or dilute mixtures required for NO control. This very dilute 
combustion required, and was understood by the inventor to require, a very 
significant tightening of statistical variations in the engine: sources of 
cycle-to-cycle air-fuel variation had to be much reduced (statistical 
variations from cylinder-to-cylinder had to be very much reduced; and 
microscale or small volume mixture variations inside the cylinder at 
combustion time had to be reduced). In addition, adequately fast stable 
combustion required control of the mixture motions inside the cylinder to 
produce intense enough turbulence for fast flame speeds with dilute 
mixtures. He has achieved the desired combustion stability by achieving 
very complete statistical uniformity of fuel, air and residuals inside the 
combustion chamber with the uniformity maintained from cycle to cycle and 
from cylinder to cylinder, and by maintaining a controllable and high 
turbulence level at combustion time so as to achieve rapid and uniform 
flame speeds. The statistical uniformity sought and achieved was and is 
much beyond the present engine art statistical uniformity, and represents 
levels of statistical uniformity well beyond levels held to be either 
necessary or desirable by authorities in automotive engineering. 
2. The trade-off between engine smoothness and peak power output has been 
recognized since before the first world war, and this trade-off is usually 
thought to be inevitable. Basically, the power output of a well designed 
engine, once its mixture has been richened to the maximum output air-fuel 
ratio, is limited by its volumetric efficiency. (An engine's ability to 
take in fuel-air mixture mass limits its power output). To get high 
volumetric efficiency, one must minimize the flow resistance of the 
passages through which charge must flow to the cylinder, so as to maximize 
the mass of mixture inducted. This entails large flow cross sections in 
passages, smooth transitions, large intake valves with large valve lifts, 
and a camshaft designed so that valves are well open at times when piston 
speeds are substantial. These modifications tend to make an engine run 
badly under the very low load conditions characteristic of conventional 
engine operation, particularly for intown driving. Under these low speed, 
low load conditions, valve overlap causes excessive charge dilution; and 
the low velocities of the mixture into the cylinder produce bad mixing and 
low turbulent flame speeds. Excessive dilution, bad mixing, and low 
combustion turbulence all tend to produce marginal combustion. 
Consequently, everyone expects racing engines to idle badly and run roughly 
under normal driving conditions. The conventional remedy for the bad 
driveability of high output type engines has been to restrict valve lifts, 
port and manifold passages, and cam timings; these modifications 
invariably restrict peak power to improve smoothness. Under conventional 
market conditions, the compromise between smoothness and peak output is 
always an uneasy one, because power and smoothness both have substantial 
market value. The wide speed-load range of engine operation makes the 
problem particularly difficult, and it is generally found that with the 
maximum tolerable fixed intake restriction, idle and low speed combustion 
stability is still much less than the stability which would be possible 
with better low speed mixing and turbulence. 
For low speeds, very restrictive intake flow passages are needed; for high 
speeds, open passages are needed, and the engine must operate under both 
low and high speed conditions. Within a fixed intake passage geometry 
format, no way out of this dilemma has been found. 
The intrinsic need for a power versus smoothness trade-off goes away if the 
intake flow passage is variable, for then the intake flow sections can be 
restrictive when high turbulence is wanted and open when high flow 
capacity is wanted; as engines are operated, the two are never wanted 
simultaneously. This variation has been achieved previously by varying 
intake valve lifts (Stivender INTAKE VALVE THROTTLING--A SONIC THROTTLING 
INTAKE VALVE ENGINE, (SAE TRANSACTIONS), Vol. 77, 1968). Intake valve 
throttling is conceptually straightforward, but involves great practical 
difficulties. Other and more producible forms of variable intake passage 
section involve very difficult and even treacherous fluid mechanical 
questions, which have occupied the inventor since the filing of the 
original case. 
3. The detailed fluid mechanics inside an engine is sufficiently 
complicated that it is held to be unreasonable to think about it or 
manipulate it in great detail. The flows are turbulent non-equilibrium 
structured flows of significant complexity both in the intake port and 
inside the combustion chamber. The concept of a structured turbulent flow 
appears to be conceptually very difficult, perhaps because turbulence is 
generally taught in exclusively statistical terms. The inventor has never 
encountered an automotive engineer who was not significantly uncomfortable 
about the very concept of flow inside a cylinder as a structured three 
dimensional flow "dance" or trajectory with random turbulent fluctuations 
superimposed on the mean flow pattern. Understanding and manipulation of 
the turbulent flow structures in engines, manipulating both flow patterns 
and velocities, is central to the current invention, and constitutes a 
significant advance beyond conventional fluid mechanical control in 
engines. 
At the time of filing the original application in February of 1975, the 
inventor had a very clear understanding of points 1 and 2 and a rather 
complete conceptual understanding of point 3 above. However, the details 
of the fluid mechanics required to produce adequately stable structures 
and controllable flows were inadequately understood. The history of the 
development of the present invention is relevant here. 
In late 1974, the inventor had the idea that continuously variable 
structured flow and turbulence in combination with very dilute mixtures 
would produce smooth and efficient combustion in combination with very low 
NO and CO emissions. At this time, a number of tests with port 
restrictions having identical flow discharge coefficients were conducted 
on a Ford 240 CID 6 cylinder truck engine. The tests clearly demonstrated 
that engine performance was much better with the kinetic energy past the 
port restriction relatively organized than with the same flow energy 
poured into turbulence, which apparently decayed too rapidly to complete 
mixing and to assist combustion. The inventor was relatively well skilled 
in the fluid mechanical art known as fluidics, which will be discussed 
below. To minimize the flow energy decay between the variable port 
restriction and the valve inlet, the inventor utilized port restricting 
means having the openings adjacent to one of the port walls (the Ford 
ports were roughly rectangular), so as to produce a Coanda wall attached 
stream flow. For this particular port, the side opening restrictions 
produced much better combustion than the top or bottom opening port 
restrictions; and it was not until much later that the inventor discovered 
that with proper port shaping, the top opening and bottom opening 
restriction designs could be made much more fluidically stable and 
effective than the side opening restrictions used in the initial 
experiments. 
Initial results with the side opening port restrictions tested at this time 
were outstandingly good. It was quickly determined that using the fluidic 
effect of wall attachment (Coanda effect) enough of the flow energy past 
the variable restriction could be maintained during the induction and 
compression process to significantly improve mixing and combustion. The 
lean limit of engine operation for this homogeneous charge engine was 
extended to the very lean ratio of 26:1 on gasoline, and engine EGR 
tolerance with a mixture leaner than stoichiometric was as much as 40% by 
mass. The variable restrictions significantly increased flame speeds for 
any set mixture quality. Engine smoothness and cold start characteristic 
also improved markedly. 
The dramatic improvement in the engine's performance at cold start was an 
initial surprise, but was easy to understand. Because of the port 
restriction, the exhaust gas blown back into the intake manifold during 
the value overlap period was a turbulent structured flow which greatly 
assisted fuel evaporation after the first engine revolution. Clearly the 
port restrictions permitted increasing valve overlap, both because dilute 
combustion was improved and because the port restriction would serve to 
limit the quantity of exhaust gases blown back into the intake manifold 
during the overlap period. 
Both theory and experiment, therefore, made the advantages of the 
inventor's variable port restriction invention very clear. Before filing 
the original case, the inventor already knew to attack the intake port 
design problem using the methods of fluidics and particularly the Coanda 
stream wall attachment effect, and had gotten good results. (For a 
discussion of fluidics, a field where flow control devices for information 
handling are made using momentum, wall attachment, turbulence decay, and 
vorticity effect see FLUIDICS, COMPONENTS AND CIRCUITS, by K. Foster and 
G. A. Parker, Wiley-Interscience, 1970). During this same time period the 
inventor and his associates put a variable port restriction modified 
engine into a Ford Maverick vehicle, under an investor dictated time 
schedule which was unreasonably fast. Largely because of these time 
pressures, the vehicle was not well worked out or calibrated. Nonetheless, 
the vehicle did operate over the EPA CVS-hot cycle with good fuel economy 
and NO.sub.x emissions approaching 0.4 grams per mile. 
However, after filing the original case, the inventor found that the swirl 
ports he there disclosed were fluid mechanically unstable, and that 
relatively small and seemingly sensible modifications of port shape 
greatly reduced the ports' measured effect on the engine combustion 
process. After doing considerable work in this fluid mechanical area, and 
trying a number of port shapes, the inventor found that turbulent decay of 
flow energy into vortex scales too small to be of use for combustion 
happened very, very quickly with most configurations tested, He further 
found that some designs that produced relatively high fluidic efficiencies 
(relatively high conversion of flow energy past the port restriction into 
forms useful for mixing and combustion) were geometrically unstable to the 
point that only quite small changes in passage shape produced very large 
differences in performance. Sometimes the changes amounted to more than a 
factor of two degradation of flame speeds, with corresponding diminutions 
in flame stability limits. 
These fluidic performance problems were crucial to the practical value of 
the variable port restriction approach. Even given perfect mixing, 
adequate performance with the dilute mixtures required for low NO 
emissions requires enough turbulence at combustion time for a reasonably 
fast burn. To shift the optimal fuel economy point toward the very lean 
mixtures required for low NO requires fast burns, because the optimal 
point is determined by the point when the thermodynamic time losses due to 
slow combustion outweigh the inherent fuel-air cycle thermodynamic 
advantages of further enleanment (or further dilution). The time losses 
due to slow combustion increase almost exactly as the 1.9 power of the 
combustion duration. Fast combustion is needed, and flame speeds are 
proportional to the turbulence intensity present at the time of 
combustion. (The relation between flame speed and turbulence has been 
strongly suspected in IC engines for at least fifty years, and was 
recently confirmed experimentally by David Lancaster, Roger Krieger, 
Spencer Sorenson, and William Hull, in "Effects of Turbulence on 
Spark-Ignition Engine Combustion", SAE paper 760160). Details of the fluid 
mechanics required to get high turbulence at ignition time will be 
discussed below. However, the value of the fluidic ports clearly hinges on 
the amount of mixing and turbulence which can be gotten per unit pumping 
work across the variable port restriction. 
In addition to prototype fluidic efficiency, the producibility of the port 
geometry is a critical commercial question. Engines are mass produced, and 
if port passage shapes must be controlled too tightly, conventional mass 
production techniques for the engine heads may not be applicable. If the 
flow becomes unstable after the deposition of thin deposits, the design is 
also impractical. It is also clear that variable port geometries involving 
many parts, close tolerances, or complex control linkages in the ports are 
undesirable. 
For these reasons, the inventor spent the bulk of the year 1976 at the 
Internal Combustion Engine Research Laboratory of the Department of 
Mechanical Engineering at The University of Wisconsin, working to develop 
a fluidically efficient variable port geometry which was also 
geometrically robust and structurally simple. He attacked the problem 
using steady state flow setups which investigated the flow using hot wire 
aenemometry and pitot tubes. The work yielded significant understanding of 
the fluid mechanics inside fluid ports and inside engine combustion 
chambers, and produced several fluidic port designs with excellent and 
robust performance. 
Because of the very complex fluid mechanics involved, men of authority in 
automotive engineering refused to concede the worth of the fluidic ports 
on the basis of these steady flow aerodynamic tests, and insisted that the 
only way the worth of the fluidic ports could ever be determined was via 
extensive and carefully conducted engine tests. 
Partly because of the potential commercial worth of the fluidic ports, and 
partially because the inventor was maintaining that wider flame limits and 
lower emissions were possible from homogeneous engine combustion than were 
considered reasonable, the engine test setup required to validate the 
performance of the fluidic ports was built with unusual care, and the fuel 
metering, air metering, and torque reading apparatus were constructed to 
significantly higher standards of accuracy than are commonly required for 
engine research. The inventor's test setup was very carefully checked not 
only by members of the Internal Combustion Engine Research Lab but also by 
outside authorities in automotive engineering, and was deemed to be 
accurate. Tests were conducted under the very close supervision of 
Professors P. S. Myers and O. A. Uyehara, who were most careful to see 
that the inventor did not draw overoptimistic or otherwise unwarranted 
conclusions from his work. Initially, Myers, Uyehara, and other ranking 
automotive engineers were by no means convinced that the good results the 
inventor expected were possible. Myers and Uyehara have outstanding 
reputations in automotive engineering and they were very well oriented 
with respect to the inventor's fluidic port results. These men did not 
consider it by any means obvious, certain, or self-evident that the 
fluidic steady state performance of the inventor's ports would be produced 
under the dynamic conditions characteristic of engine operation, nor did 
they find it self-evident that proper fluidic performance of the ports 
would permit the efficient and stable ultralean combustion expected by the 
inventor. These men doubted that the very dilute mixtures required for 
ultralow NO.sub.x with homogeneous combination could be burned with 
adequate stability and speed, no matter how tight mixing statistics were 
and no matter how well controlled turbulence levels were. On the basis of 
the evidence then available, in light of traditional automotive 
engineering standards of evidence and logic, these doubts were entirely 
rational. 
Automotive engineering requires, for clear commercial reasons, very high 
standards of proof, particularly with respect to technology concerned with 
so sensitive a subject as emissions. The present application is somewhat 
detailed and lengthy because it contains the information required to teach 
skilled automotive engineers, who are not commonly acquainted with 
fluidics and have difficulty visualizing and thinking about structured 
flows, how to understand, make, and use fluidic ports in combination with 
lean mixtures to produce very efficient and ultralow NO.sub.x output 
engines with improved peak power and excellent driveability. Also, the 
application is lengthy because it contains the detailed experimental 
results which the automotive engineering profession deems necessary (if 
not, perhaps, sufficient) to establish the technical worth of the 
inventor's ultralean fluidic port invention. In addition to the inventor's 
own experimental work, other experimental work is discussed in order to 
explain the invention's performance and more clearly orient the invention 
in the context of prior automotive engineering knowledge.

DETAILED DESCRIPTION 
See FIG. 1, which is a cross section side view of the fluidically simplest 
of the inventor's fluidically efficient variable restriction ports. In the 
port of FIG. 1, the variable restriction opens to form a wall attached 
stream flow along the top of the port surface, continuing smoothly out of 
the valve opening to provide a coherent and highly structured flow for 
controlled and rapid mixing and fast flame speeds in the engine. Port 
structure 1 includes swinging throttle plate 2 mounted on rotating shaft 3 
located at the floor of the port in such a manner that there is little 
flow leakage between shaft 3 and the port floor juncture at 4. The port is 
shaped so that the sides of the throttle plate 2 (not shown) seal 
relatively well along the sides of the port as the throttle swings from 
open to closed position; thus the combination of the top opening throttle 
and the port forms an airflow opening of variable area at the top of the 
port. Because of well known fluid inertia effects, flow streamlines past 
the restriction converge to form the well known vena contracta at 5. At 
this vena contracta, point 5, flow velocity is very near the isentropic 
flow velocity corresponding to the pressure drop across the variable port 
restriction. Stream 6 is attached to the top surface 7 of the port 1 by 
means of the fluid principle of wall attachment (Coanda effect). (For a 
detailed explanation of the Coanda effect, see pages 131-139 of FLUIDICS, 
COMPONENTS AND CIRCUITS, K. Foster and G. A. Parker, Wiley Interscience, 
1970). 
The principle of wall attachment is important enough that is must be 
described in some detail here. Basically, a jet entrains flow on both 
sides. If the jet is near a wall, fluid entrainment generates a reduced 
pressure on the wall side with respect to the outside of the jet flow. 
Because of the pressure difference between the wall side and the outside 
of the jet, the jet flow path curves toward the wall (the jet is sucked 
towards the wall). As the jet bends toward the wall, the wall pressure 
becomes smaller, the suction stronger, and in consequence the jet attaches 
to the wall. This attachment effect is utilized in a number of important 
digital fluidic logic circuits, for example those invented by Raymond 
Warren et al., at Harry Diamond Laboratories. The Coanda, or wall 
attachment, effect is of great importance to the design of the inventor's 
fluidically efficient variable ports for several reasons. First, a wall 
attached stream spreads much more slowly than a free jet, so that the 
kinetic energy in the flow is dissipated much more slowly for a wall 
attached stream. The slower spreading permits the variable restriction to 
be placed at a much greater distance from the intake valve opening than 
would otherwise be possible. Second, a wall attached stream will follow a 
wall curvature when momentum effects on the stream would otherwise cause 
separation, and this curve following property facilitates stream control. 
Third, for the Reynolds numbers characteristic of intake ports, proper 
passage curvature control permits the jet flow to be controlled so that 
the wall attached stream can be made to detach from the wall cleanly at a 
specified point and moving in a specified direction. 
Referring again to FIG. 1, the onrushing attached jet 6 rushed past the 
specially faired port surface in the vicinity of the valve guide at 8 and 
continues downward to the valve seating surface, where an intermediate 
curvature at 9 permits the Coanda wall attachment effect to curve the jet 
to be tangential with the valve seating surface. Thus the jet leaves the 
passage surface and rushes into the combustion chamber in the form of a 
still quite coherent high velocity sheet of flow directed to an angle from 
the valve stem of approximately 45.degree. and with the flow occurring 
virtually exclusively over less than 180.degree. of the valve face. This 
flow is oriented to produce both swirl about the engine cylinder axis and 
a useful axial velocity component for structured flow mixing parallel to 
the axis of the engine cylinder (not shown). 
Flow velocity vectors 10 show a characteristic velocity profile at the 
point of flow introduction into the engine combustion chamber. Note that 
the high velocity flow does not extend out to the intake valve, but is 
thinner. This is important: in order for flow velocity for a fixed volume 
flow to be increased, the cross sectional area across which the flow 
occurs must be decreased. Fluidic variable restriction ports permit this 
without changing valve lifts and without constraining WOT engine 
performance. 
FIGS. 2A and 2B show the velocity distribution around the valve of the port 
design of FIG. 1, with broken lines in FIG. 2B to show the axis of the 
port passage. Shaping of the upper surface of this port shape involves a 
valve guide entry which is difficult to machine and a passage shape 
somewhat constrained in flow capacity when compared to a racing port 
design. Other port designs to be discussed later are better for flow and 
more practical. However, the port shape shown in FIG. 1 is fluidically the 
simplest of the variable restriction fluidically efficient ports. The port 
is characteristic of the class of fluidic ports in important ways. 
For this port, as with the other fluidically efficient variable ports, the 
wall attached stream energy decay is sufficiently slow that the 
restriction can be placed a substantial distance from the valve seating 
surface from which the flow rushes into the cylinder. The restriction can 
be, for example, more than 15 cm. from the valve. Also, the variable 
restriction could be arranged to swing inwardly toward the valve rather 
than swinging outwardly away from the valve. Also, the variable 
restriction need not be a swinging throttle, but could be a guillotine 
type sliding throttle. 
Most importantly, the fluidic port design of FIG. 1 characterizes its class 
in that its permits, by variation of the setting of its variable 
restriction, variations in the flow structure and flow energy into the 
engine cylinder. When restricted, the port design produces a coherent and 
strong flow pattern which will dominate the detailed structured flow 
patterns, turbulence, and flame speed of the engine. With a fluidic port 
which produces this controlled and structured intake flow pattern, it is 
possible to operate a homogeneous charge engine with excellent efficiency, 
excellent smoothness, and high peak power. For normal driving conditions 
the engine can operate with such dilute mixtures that the engine produces 
only trace values of NO.sub.x, and with combustion characteristics such 
that the calibration for minimal NO.sub.x emissions produces optimal or 
near optimal economy. 
This low NO.sub.x performance is very surprising from the point of view of 
conventional automotive engineering doctrine, and goes against great 
bodies of previous engineering work. The reasons for the excellent 
performance of a properly structured and high velocity intake flow involve 
central matters in the interaction of fluid mechanics with mixing and 
turbulence, and in the interaction of mixing and turbulence with the 
physical chemistry of flame stability and NO formation in engines. Many of 
these matters relating to flame limits and mixing are not well understood 
in automotive engineering, and are crucial to an understanding of the 
present invention. Consequently, included here is a detailed discussion of 
the chemical, statistical, and fluid mechanical matters which make the 
excellent performance of the present invention possible. Wherever 
possible, references supporting the central points have been taken from 
the automotive engineering literature in order to maximize the credibility 
of the explanations and to put the present invention in its context with 
respect to other automotive engineering work. 
THE PROCESS OF NO.sub.x FORMATION IN ENGINES: 
EFFECTS OF EQUIVALENCE RATIO AND MIXING 
First, an understanding of the mechanism by which NO.sub.x is formed in 
engines is important in order to understand the effects of stoichiometry 
and postflame gas heat transfer relations on engine NO.sub.x output. 
FIGS. 3, 4, and 5 are taken from "NITRIC OXIDE EMISSIONS FROM STRATIFIED 
CHARGE ENGINES: PREDICTION AND CONTROL," by Paul N. Blumberg, Scientific 
Research Staff, Ford Motor Company, which was published in March of 1973. 
The chemical processes which produce NO.sub.x in engines are different in 
kind from the processes which generate CO and HC emissions. CO and HC are 
the result of incomplete combustion (due to local insufficiencies of 
temperature or of oxygen in the fuel-air mixture). Nitric oxide, on the 
other hand, is the product of an endothermic reaction at very high 
temperatures which bonds atmospheric oxygen to atmospheric nitrogen to 
form NO. 
The NO forming reaction is an endothermic reaction which partly depends on 
the amount of oxygen which is left over after burning the fuel, so that 
the equivalence ratio of the mixture has an effect over and beyond the 
effect produced by temperature differences. However, NO formation rates 
even at peak temperature are slow enough that NO concentrations in any 
element of gas are only incidentally related to chemical equilibrium, and 
temperature effects tend to greatly overshadow concentration effects. NO 
formation is dominated by kinetic reaction rates. These reaction rates are 
very sensitive to temperature. Relatively small decrements in mixture 
temperature can produce large reductions in NO formation rates, and 
therefore produce large reductions in NO output concentrations. Because 
chemical kinetics of NO formation is so slow, NO formation is not a flame 
chemistry problem, per se. The NO is formed in the post flame gases. 
FIG. 3 (which is also Blumberg's FIG. 3) shows the effect of equivalence 
ratio and EGR on NO production for a homogeneous charge engine, and is 
based on an important chemical kinetics model of the NO forming process. 
For efficient engine operation and CO and HC control, the interesting 
mixtures are at, or leaner than, an equivalence ratio of 1.0. 
(Relationships between equivalence ratio and air-fuel ratio are clearly 
shown on the ordinate of the graph.) For these lean mixtures, NO output 
levels are dominated by mixture variations which effect flame and 
postflame temperatures. 
For example, on the E.G.R.=0% line, leaning the mixture from an equivalence 
ratio of about 0.96 produces very sharp decreases in NO output, even 
though available oxygen increases as the mixture is leaned out, because 
the leaner mixtures have less internal energy per unit mass and 
consequently have lower peak flame temperatures and lower temperatures 
during the power stroke of the engine. As the mixture is leaned out from 
roughly stoichiometric, peak temperatures decrease because the excess air 
serves as a diluent, and the thermal capacitance of this excess air lowers 
peak temperatures. (This dilution reduces temperatures and, therefore, 
cuts the dissociation of chemical species which otherwise produces a 
thermodynamic loss in the engine cycle, so that the diluent actually 
improves the fuel-air cycle efficiency of the engine). 
Exhaust gas recirculation (EGR) serves as a diluent and lowers temperatures 
also because of its "thermal capacitance" (thermal capacitance is defined 
as mass times specific heat), and except for relatively small variations 
due to oxygen concentration changes, the effects of air and EGR as 
diluents are similar for like thermal capacitances. (Flame chemistry and 
inlet temperature relationships are such that most engines will tolerate 
more dilution via EGR than with simple enleanment.) It follows that 
combinations of excess air and residual gas can be used as diluents as 
well as either alone. In this specification dilute lean combustion will be 
defined as combustion of a mixture leaner than stoichiometric where the 
combined effect of the thermal capacitances of the excess air and the 
residual gases is such that peak flame temperatures and temperature 
trajectories for the postflame gases are depressed to the point where very 
little NO.sub.x is produced. 
Even for a homogeneous charge engine, the formation of NO is not uniform 
throughout the charge because the NO formation is an integrated effect 
from the temperature-pressure-time history of the successive elements of 
the charge to burn. The NO.sub.x output of an engine will depend to a 
significant extent on the exact time-temperature-pressure path (kinetic 
trajectory) of its postflame gas elements, which will depend significantly 
on the mixing and flow structure of the engine and on spark timing. 
FIGS. 4 and 5 (which are also FIGS. 4 and 5 of Blumberg, op. cit.) give as 
good a description as any two graphs might of the interrelationships of 
temperature, time, and pressure in the formation of NO.sub.x in an engine. 
The curves are calculated on the basis of no heat transfer between 
successive elements of mixture to burn, so that adiabatic relationships 
can calculate the temperature-time-pressure trajectories which will 
determine NO formation. The graphs will make clear how sensitive NO 
formation can be to mixing and heat transfer between successive elements 
of the postflame gases. 
FIG. 4 shows the temperature versus crank angle trajectory for the 
postflame gases from the first element of (homogeneous) mixture to burn, 
the middle element of the mixture to burn, and the last element of the 
mixture to burn. FIG. 5 shows the NO concentrations formed in these 
elements versus crank angle (the lines of interest are the rate calculated 
concentrations, which represent the concentrations of NO in the postflame 
elements as a result of past temperature-pressure-time conditions). Engine 
operating conditions are printed on the figures. 
In FIG. 4, look first at the temperature trace of the first element (which 
begins when the flame passes through this element 10.degree. before top 
dead center crank angle). After the mixture burns for this first element, 
successive elements burn in the chamber and increase chamber pressure, 
adiabatically recompressing these first postflame gases to burn. The 
consequence is that the temperature of the postflame gases from the first 
element are higher than the temperature of the initial flame for about 
seventy crank degrees after the flame has passed. Note also that the peak 
gas temperature for this element is achieved more than thirty crank 
degrees after the flame has passed in this calculation. 
Temperature trajectories for the middle element to burn and the last 
element to burn are calculated in the same way and also plotted in FIG. 4. 
FIG. 5 shows NO versus crankangle for the temperature-pressure-time 
trajectories plotted in FIG. 4. The difference between the NO 
concentration in the first element to burn (5600 ppm) and the NO 
concentration in the last element to burn (1000 ppm) is dramatic. Lower 
temperatures and lower high temperature residence times both greatly 
effect NO outputs. 
Although the adiabatic element assumption of Blumberg's model is never 
entirely valid, the graphs show the relationships between 
temperature-pressure-time trajectories and NO concentrations and should 
help clarify the following important points: 
1. Variations in heat transfer between successive elements of mixture to 
burn can change No outputs by changing temperature-pressure-time 
trajectories. Since No formation increases with temperature at much more 
than a linear rate, any such mixing will tend to reduce NO outputs. The 
detailed effect will depend on the details of the mixing process due to 
flow structure which determines kinetic trajectories. 
2. NO is formed on an element by element basis, and this means that the NO 
output from an engine operating on a specific fuel-air-residual mixture 
will depend on the degree of stratification or homogeneity which exists in 
the cylinder. From FIG. 3 it should be clear that, particularly for lean 
mixtures, homogeneity is much preferable to stratification both from the 
point of view of NO formation and from the point of view of efficiency. 
This point was made in Dr. Blumberg's paper from which FIGS. 3, 4, and 5 
come. 
The importance of temperature-time-pressure trajectories shown in the 
above-discussed figures also explains the importance of spark timing and 
flame speed on the NO results. Dramatic examples of spark timing effects 
on NO formation will be shown for the ultralean variable structured flow 
and turbulence engine of the present invention. 
It should be clear from a consideration of FIGS. 3, 4, and 5 that mixing is 
very important in an engine if very low NO ouputs are required. For dilute 
mixtures, any heterogeneity can drastically increase NO levels. Another 
reason mixing is important concerns flame stability and smooth operation 
of the engine. With the relatively low quality mixing of prior art 
engines, satisfactory combustion with the very dilute mixtures required 
for low NO is almost impossible because of statistical variations in 
mixture strength from cylinder to cylinder, from cycle to cycle and within 
the cylinder. It will be shown that flow within an engine is invariably a 
significantly structured (nonrandom) flow pattern, on the basis of well 
established automotive engineering data and the inventor's data. 
Automotive engineers commonly do not understand and cannot visualize the 
important effects flow structure has on the actual mixing phenomena in an 
engine cylinder. Since the operability of the present engine hinges on 
these effects, they must be explained. 
TIGHT STATISTICS PERMIT STABLE OPERATION WITH MORE DILUTE MIXTURES 
FIG. 6 gives a graphical explanation of how improved mixing within the 
combustion chamber can widen the equivalence ratio or dilution limits 
which permit stable combustion inside an engine. This statistical argument 
is at the very core of the conceptual background of the present invention 
but it involves statistical arguments which are difficult to visualize. 
FIG. 6 is an illustration of a statistical numerical example of the 
argument, and is intended to clarify an argument which is otherwise hard 
to follow. FIG. 6 illustrates variations in a hypothetical engine where 
the gross air-fuel-residual ratios from cycle to cycle are invariant but 
where the mixing inside the cylinder is less than perfect. 
Experimentally, the inventor has shown that homogeneous engine operation 
with mixtures as lean at 0.55 equivalence ratio is possible, but for the 
purpose of the graphical example of FIG. 6, suppose that if the mixture 
within the spark plug gap is leaner than 0.55 equivalence ratio at the 
time of sparking, misfire will infallably occur (this is a worthwhile 
oversimplification for the present purpose). As engines currently operate, 
the transition between steady firing and misfire is not abrupt--misfire is 
defined statistically, with a misfire frequency above a certain frequency 
scored as unacceptable. A 1% misfire rate is often used to define the 
misfire limit. (For detailed discussions of these matters, see "Lean 
Combustion and the Misfire Limit in Spark Ignition Engines" SAE paper 
741055 and "What Limits Lean Operation in Spark Ignition Engines--Flame 
Initiation on Propagation" SAE paper 760760, both by Ather A. Quader of 
General Motors Research Laboratory.) 
(Here, in addition to the misfire limit, a partial burn limit is defined; 
but partial burn need not concern us at this point in the discussion.) It 
should be clear that the mixing quality of the mixture, as measured by the 
statistical distribution of small charge element equivalence ratios about 
the cylinder mean equivalence ratio, will affect the misfire rate for any 
given cylinder mean equivalence ratio. Very much leaner overall 
equivalence ratio operation will be tolerable vis-a-vis misfire if the 
statistical uniformity of the mixture is much increased. 
See FIG. 6. Curves A, B, and C are plotted as gaussian distributions, and 
the plots are for such small sample volumes (for example, cubes 1 mm/side, 
which is about the volume in the spark plug gap at ignition time) that the 
distributions can be taken as continuous distributions. The integrals 
under curves A, B, and C are equal. Curve A has a mean equivalence ratio 
of 0.75 equivalence ratio, but has a standard deviation of 0.1 equivalence 
ratio for its distribution. Under the assumptions of FIG. 5, this mixture 
distribution A will misfire in the engine 2.25% of the time: mixture 
quality A can be said to be at its lean misfire limit at a ratio somewhat 
richer than 0.75 stoichemetric. The standard deviation of the mixture 
plotted on curve B is half the standard deviation of curve A, or 0.05 E.R. 
A 2.5% misfire rate for distribution quality B occurs at an overall 
equivalence ratio of 0.65 stoichemetric, and so an engine with mixing such 
as that shown for curve B will have a misfire limit slightly richer than 
0.65 equivalence ratio. 
Curve C is shown with a standard deviation of 0.01 E.R., and with the 
overall equivalence ratio of the mixture at 0.585. This mixture is much 
leaner than that of curve A or curve B, but because of its tight mixing 
statistics, a mixture leaner than 0.55 E.R. will occur in the spark gap 
less than one tenth of one percent of the time. Mixture distribution C, 
with its mean at 0.585 stoichemetric, will have a misfire rate twenty-five 
times less than the misfire rate distribution A even though A has a mean 
equivalence ratio of 0.75 E.R., and distribution C will also have a 
misfire rate only one twenty-fifth as great as that of distribution B with 
its ratio of 0.65 stoichemetric. Better mixing (tighter statistics) than 
that shown in curve C would permit the misfire limit to be approached even 
more closely. Tightening mixture distributions in the cylinder permits a 
much closer approach to the ultimate physical misfire limits in an 
operating engine. Operation with these much more dilute mixtures produces 
dramatically reduced NO emissions, and if turbulence is right, engine 
efficiency is increased simultaneously. 
The argument of FIG. 6 was addressed to an engine where cycle-to-cycle and 
cylinder-to-cylinder statistical variations in mixture quality were 
negligible. However, an exactly symmetric argument, easily visualized by 
reference to FIG. 6, exists showing that statistical variations of 
cycle-to-cycle or cylinder-to-cylinder mixture quality will limit the 
mixture dilution tolerable with respect to engine operation even if 
in-cylinder mixing were perfect on each cycle. Tight control of 
cylinder-to-cylinder, cycle-to-cycle, and microscale mixing statistics are 
all necessary conditions (none of which is sufficient in itself) to stable 
operation with the very dilute mixtures required for efficient low 
NO.sub.x engine operation. 
The statistical argument illustrated in FIG. 6 is very fundamental to the 
function of the present invention. Although the illustration is 
oversimplified (for example, air-fuel-residual fraction proportions rather 
than just air-fuel ratios effect misfire), the oversimplifications do not 
change the principle. For best torque spark timings, misfire occurs 
primarily because of inadequate combinations of air-fuel-residual adjacent 
the spark gap at sparking time, and as the spatial distribution of 
chemical species in the cylinder becomes more and more uniform, the 
overall chemical ratios in the cylinder can approach the limiting 
conditions more and more closely. The inventor's data shows that this 
extension of combustion limits permits NO.sub.x outputs far below any 
proposed NO.sub.x standard to be achieved with excellent fuel economy. 
A consideration of FIGS. 3, 4, 5, and 6 should make it clear that mixing of 
fuel-air and residual are important for engine operation and NO levels, 
and also that flow structures which effect heat transfer relations between 
adjacent elements of postflame gases can effect NO.sub.x outputs. These 
points are significant to the present invention, which produces rapid and 
controlled mixing which permits very dilute combustion and produces very 
efficient and low NO.sub.x operation with dilute mixtures. 
RELEVANCE OF MIXING 
The fluid mechanics of mixing in the engine must be understood for the 
function of the present invention to be understood. 
Turbulent fluid mechanics as it is conventionally taught is an extremely 
forbidding subject, and the study of mixing in turbulent flows is a 
relatively complicated part of turbulence theory. It is not unfair to say 
that the in-cylinder fluid mechanics governing mixing inside cylinders is 
dismissed by the bulk of the automotive engineering profession as too 
complicated to consider. It is also fair to say that automotive engineers 
do not understand the way flow structure and turbulence interact in 
engines to produce mixing. Particularly, automotive engineers commonly do 
not understand the fact that the reproducible flow structure, which is the 
same from cycle to cycle and on which the random fluctuations of 
turbulence is superimposed, is at least as important as turbulence in 
eliminating any vestiges of charge stratification in the cylinder. 
When the inventor commenced work on improving mixing in cylinders many 
years ago, he was convinced on the basis of detailed physical arguments 
and analogies with other mixing systems that mixing in conventional spark 
fired engines was inadequate for efficient dilute combustion. These 
arguments involved order-of-magnitude calculations which were admittedly 
inaccurate, but estimated magnitudes were persuasively large. These 
arguments were not obvious to a conventional automotive engineer, and 
indeed would have been considered unacceptable. Order-of-magnitude 
arguments are not considered an adequate basis for action in automotive 
engineering. 
However, during the time that the inventor has been perfecting his 
invention, some very detailed measurements on turbulence inside engines 
have been undertaken using hot wire aenemometry and sophisticated data 
processing techniques. This data has made it possible to show 
mathematically how slow turbulent mixing per se is in engines, and also to 
show that structured turbulent flows can and do exist inside the 
combustion chamber. 
MODELLING OF TURBULENT DIFFUSION IN CONVENTIONAL ENGINES 
Probably the best work in this field of in-cylinder turbulence and flow 
measuring has been done at the General Motors Research Laboratories, and 
is reported in two SAE papers, "Effects of Engine Variables on Turbulence 
in a Spark-Ignition Engine" by David R. Lancaster, SAE paper 760159, and 
"Effects of Turbulence on Spark-Ignition Engine Combustion" by David. R. 
Lancaster, Roger B. Krieger, Spencer C. Sorenson and William L. Hull, SAE 
paper 760160. This work was done on a standard split head CFR research 
engine. Another piece of work which produces very relevant evidence with 
respect to in-cylinder mixing rates was done on a basically identical 
split head CFR, permitting direct confirmation of mixing quality from the 
turbulence levels measured by Lancaster and associates. This work is 
"Cloud Combustion A Study of Performance and Emissions," an unpublished 
M.A. thesis in Chemical Engineering done by George A. Oliver at the 
California Institute of Technology, June 4, 1973. 
With data from Lancaster's SAE paper 760159, one can show quite strikingly 
that mixing due to turbulent diffusion alone is very inadequate to produce 
homogeneous mixtures in the cylinder. In light of the argument related to 
FIG. 6, this is a crucially important point. Using the reference cited by 
Lancaster for calculating turbulent diffusion coefficients (Hinze, J. O. 
Turbulence McGraw Hill, New York, 1959 p. 361), we find that the turbulent 
diffusion coefficient E equals 
E.sub.P.sbsb. =.infin. =.mu.'.LAMBDA..sub.e 
where .mu.'=root mean square turbulent fluctuating velocity 
.LAMBDA..sub.L =integral turbulence scale 4.times.microscale spatial 
length 
Using the data of Lancaster's SAE paper 760159 cited on Page 13, the 
turbulent diffusion coefficients characteristic of his engine operated 
with a conventional inlet valve at 1200 RPM is 3.53 m.sup.2 /sec. From the 
same source, the turbulent diffusion coefficient characteristic of 1200 
RPM operation with a tangentially oriented shrouded valve is 
15.2.times.10.sup.-3 m.sup.2 /sec. At 1200 RPM there are 50 milliseconds 
for the intake and compression stroke. 
The CFR engine combustion chamber is a disk with varying thickness as the 
piston moves. For this reason, a two dimensional mixing model can be used 
to model the mixing process in the engine. Species diffusion calculations, 
which are exactly analagous to thermal diffusion calculations, are best 
modelled by computer using the finite element technique well known to the 
engineering arts. An associate of the inventor, Kenneth Kriesel, made a 
computer model of the turbulent diffusion process in the engine, assuming 
only turbulent diffusion, using the computation program of Professor Glenn 
Myers of the University of Wisconsin. 
CONVENTIONAL ENGINE TURBULENT DIFFUSION SLOW 
FIG. 7 shows the initial condition for the calculation. For reasons of 
computational simplicity, the initial conditions were set up so that all 
of the fuel was introduced in a pie shaped disk 11. The model then 
proceeded for the fifty millisecond modelling time. 
FIG. 8 shows the concentration gradients resulting using the diffusivity 
calculated from Lancaster's data for the nonshrouded valve case (this 
turbulent diffusivity should be close to the turbulent diffusivity of a 
conventional production engine operating at 1200 RPM). After the 50 
milliseconds concentration variations of .apprxeq.84:1 still persist 
inside the cylinder. 
FIG. 9 shows the concentration gradients resulting using the diffusivity 
calculated for the shrouded valve case (a diffusivity substantially higher 
than that typically seen at 1200 RPM in a production engine). For this 
case concentration variations of .apprxeq.4.6:1 still persist in the 
cylinder after the intake and compression stroke are completed. 
It is clearly true that the modelling process used is not perfect, and that 
the initial condition heterogeneity was worse than would occur in any 
engine practice. However, even changing diffusivities by factors much 
beyond any that are consistent with Lancaster's and other's measurements, 
the conclusion is still inescapable that turbulent mixing per se in 
engines is far from adequate to produce true homogeneity in the cylinder 
at ignition time. 
Turbulent diffusion, in itself, is not nearly fast enough to produce 
adequate combustion in engines, particularly when the volume of residual 
gas which must be mixed with the incoming charge is large, or if the 
mixture inducted is in some way stratified as it usually will be if liquid 
phase is present. However, certain kinds of flow structures can increase 
mixing rates by extraordinarily large factors, while other flow structures 
can have little effect on mixing. An understanding of the interaction of 
flow structure and turbulence is vital to practical mixing in engines. So 
far as the inventor is aware, this vital interaction has not been clearly 
understood previously in automotive engineering. 
FLOW STRUCTURE RADICALLY EFFECTS MIXING 
FIGS. 10, 11, 12, 13, 14, 15, and 16 show that mixing in a well defined 
statistical sense, can occur in totally laminar flows in the absence of 
diffusion, producing a close and predictable geometrical relation between 
the fluids to be mixed. The structured flow will produce a structured 
distribution of the different fluids, and in this way the mean distances 
over which turbulent and molecular diffusion will have to act can be 
dramatically reduced. For the idea for FIGS. 10-16, I am indebted to T. M. 
McKelvey (Page 299 POLYMER PROCESSING, John Wiley & Sons, 1962). 
Mathematically, it can be said the flow structure will geometrically 
transform all the points in a concentration fluid in a determinate way as 
a function of time, and is a one-to-one transform function. 
FIG. 10 shows the velocity distribution about the radius of an irrotational 
flow vortex. Irrotational vortices are among the most common flow patterns 
in nature (whirlpools, turbulent vortices, tornados, hurricanes, etc.) and 
there is evidence for the existence of a large irrotational vortex in the 
inventor's engine data. In an irrotational flow vortex, the fluid angular 
momentum, mvr, is constant at any radial distance from the center of the 
vortex. For a velocity at any radius V.sub.r, the tangential velocity 
equation is V.sub.r =r.sub.o /r V.sub.o, where r is radial distance of the 
fluid element from the vortex center, r.sub.o is the outside radius of the 
vortex, and V.sub.o is the tangential velocity characteristic of the fluid 
at radius r.sub.o. 
FIG. 11 shows an irrotational flow vortex where a line of mixant has been 
introduced instantaneously. FIG. 12 shows how the structured flow of the 
irrotational flow vortex has stretched out the mixant, assuming zero 
turbulence and zero molecular diffusivity, in the time required for the 
outside of the vortex to rotate 90.degree.. The flow stretching occurs 
because the angular velocity of an element in the vortex, varies as 
.theta..sub.r =r.sub.o .theta..sub.o /r.sup.2. It is important to notice 
that this flow stretching is important for mixing but that it does not 
involve any randomness at all. 
FIG. 13 shows an irrotational vortex analagous to the vortex of FIG. 11, 
wherein four perpendicular radial lines of mixant have been introduced 
instantaneously. FIG. 14 shows how the structured flow of the irrotational 
flow vortex has stretched out the mixant, again assuming no diffusivity, 
after the outside edge of the vortex has rotated 90.degree.. It should be 
noted that much more than 90.degree. outside rotation should be expected 
under engine conditions. 
Clearly, there are structured flows which stretch out fluid elements in 
such a way as to greatly increase interfacial area and much reduce the 
mean distances over which turbulent and molecular diffusion must act. 
However, other structured flows have little or no such effect. 
FIG. 15 shows the velocity distribution for a rigid body fluid rotation to 
show a flow structure which does not mix. Rigid body flow rotation has 
been shown to exist in a number of engines (See, for example, "Measurement 
of Air Movements in Internal Combustion Engine Cylinders," by M. Horvatin 
and A. W. Hussman in DISA INFORMATION FF 8, July 1969) and is the flow 
model generally used to model swirl in internal combustion engines. FIG. 
16 shows a radial line of mixant introduced instantaneously in the rigid 
body fluid rotation. Since the angular rotation of the fluid elements is 
the same for each radius, rotation does not produce stretching or 
redistribution of the fluid concentration structure for the rigid body 
rotation case. FIG. 16 will serve as a picture of the mixant distribution 
after any integral number of revolutions of the solid body fluid rotation, 
just as surely as a stripe on a wheel would appear the same after n 
revolutions. Clearly, solid body fluid rotation is of no use to the mixing 
process. 
Many structured flow patterns more complex than those shown exist. In 
general, those which produce large local velocity gradients in the 
cylinder will be useful for mixing in an engine. 
The interaction of flow structure and turbulence in engines is of very 
great importance if homogeneity is to be achieved. By decreasing the 
distance over which turbulent diffusion (and on a smaller scale, molecular 
diffusion) will have to act, structured flow can homogenize a fluid mass 
very quickly. Fundamentally, this is because the rate at which turbulence 
destroys heterogeneity goes as the inverse square of the distance over 
which the diffusion must act. The differential equation for turbulent 
diffusion is as follows: 
EQU N.sub.a =-D.sub.v .delta.c/.delta.s 
N.sub.a =diffusion rate at a point 
c=concentration 
s=distance (Perry's Handbook 14.4) 
D.sub.v =diffusivity 
If a flow structure cuts the distance across which diffusion must act 
threefold, the concentration gradient in the fluid increases by threefold, 
and mixing rates increase by a factor of three on this account. But the 
distance across which the diffusion must occur is also cut threefold. This 
means that the mass flow across any plane which must occur to equalize 
concentrations is also cut threefold. The mass flux effect and the 
gradient effect are multiplicitive, so that cutting the mean distance 
across which diffusion must act threefold cuts the time required for 
equalization three times threefold, or ninefold. 
If the mean distance across which diffusion must occur had been cut 
tenfold, the time of mixing to produce a set level of homogeneity would 
have been cut a hundredfold. With this inverse square mixing rate in mind, 
another look at FIGS. 10-16 should convince the reader that flow 
structures within an engine cylinder can have overwhelmingly large effects 
on the mixing quality achieved by ignition time. 
Even though an exact calculation of the convolutions of the structured flow 
and the nature of the turbulent flow field is essentially never possible 
in an engine, an understanding of this interaction between flow 
concentration stretching structured flows and turbulent mixing is of very 
great importance. Variations in flow structure from engine to engine can 
make very great differences in the mixture homogeneity of the engines. 
Many of the unpredictable "driveability" variations from combustion 
chamber to combustion chamber are probably due to variations in 
in-cylinder mixing rates due to variations in flow structure. 
STRUCTURED FLOW EXISTS IN ENGINES AND CAN HAVE DOMINANT EFFECTS ON MIXING 
Flow inside any reciprocating piston engine must involve a definite flow 
structure for basic physical reasons. All the fluids in an engine have 
inertial mass, and the velocity of the flow into the engine combustion 
chamber through the intake valve(s) is always much greater than the 
velocity of the center of mass of the gases in the cylinder. The center of 
mass of the fluid inside the cylinder moves generally at less than half of 
piston velocity. For conventional engines as presently produced, the mean 
inlet velocity is generally more than nine times piston velocity. The 
ratio of the inlet mean velocity to the in-cylinder charge center of mass 
velocity is, therefore, more than 18 to 1. In terms of momentum fluxes, 
this means that the intake flow will have to produce some sort of ordered 
swirling motion whereby the fluid momentum can be contained in the 
cylinder as the momentum of the flow is dragged down by interactions with 
the cylinder, head, and piston surfaces. The swirling motion need not be 
what automotive engineers call swirl, which is rotation of the whole 
charge about the axis of the cylinder, but may for instance involve a 
system of larger and smaller vortices and flow paths, with significant 
flows parallel to the axis as well as perpendicular to the axis of the 
cylinder. The momentum of the fluid will not instantly disappear as it 
enters the cylinder: therefore, some ordered flow inside the cylinder has 
to happen. Because the basic inertial input conditions will be nearly 
identical from cycle to cycle, the gross flow pattern in the cylinder, 
which is dominated by nonrandom inertial and pressure effects, will be 
nonrandom, too. A good way of thinking about the flow inside the cylinder 
is as a convoluted three dimensional hydrodynamic "dance" which will be 
dominated by the reproducible intake flow conditions and will, therefore, 
be much the same from cycle to cycle. Superimposed on this large scale 
hydrodynamic "dance" will be the more or less random perturbations of 
turbulence. 
The concept of a structured flow on which turbulent fluctuations are 
superimposed is an uneasy one for most automotive engineers. Fortunately, 
Lancaster in SAE paper 760159 measured the flow inside a motored engine 
with both a shrouded and unshrouded intake valve, and definitely showed a 
pronounced flow structure for both engine cases. Since the concept of a 
structured turbulent flow is central to an understanding of the function 
of the present invention, and since the concept is solidly established by 
means of the work of Lancaster et al., so that it is no longer a matter of 
conjecture, the data of SAE paper 760159 needs to be considered in detail 
here. 
The following figures are taken from Lancaster's SAE paper 760159. FIG. 17 
(Lancaster's FIG. 3) shows the layout of the hot wire aenemometer sensors 
on the CFR split head used in the experiments (shown with the shrouded 
valve installed). FIGS. 18a, 18b, and 18c (Lancaster's FIGS. 4a, 4b, and 
4c) shows the relationship between the engine cycle (18a) the hot wire 
signal of instantaneous velocities (18b); and the separation of the hot 
wire signal averaged for many cycles into a mean velocity trace and RMS 
turbulent intensity trace corresponding to turbulent fluctuations 
superimposed on the mean flow. The ensemble averaged mean velocity shown 
in FIG. 18c is the result of averaging digitized degree-by-degree 
velocities for 100 successive cycles. Any average over 100 cycles will 
strongly smooth out random fluctuations. Therefore, the very sharp peaks 
and valleys of the mean velocity curve in FIG. 18c are extremely 
conclusive evidence of a structured flow within the cylinder charge. It 
should especially be noted that the mean velocity component of the fluid 
motion is very much greater than the turbulent fluctuations superimposed 
upon it. The kinetic energy of the flow is dominated by the hydrodynamic 
"dance" described above. The hydrodynamic dance is caused by basic 
inertial physics. 
Lancaster's very commendable paper shows similarly conclusive evidence for 
structured flow from the nonshrouded intake valve characteristic of 
conventional production engines. FIG. 19 (Lancaster's FIG. 5) shows the 
ensemble averaged mean velocity and turbulence intensity characteristic of 
the nonshrouded valve under the same operating conditions as those of the 
previous figures (1500 RPM, 50% volumetric efficiency, 8.72:1 compression 
ratio). Note again that the mean velocity is much greater than the 
turbulent fluctuations imposed upon it. However, in FIG. 19, note how much 
smoother the mean velocity curve characteristic of the nonshrouded valve 
is: this indicates that the velocity gradients within the flow are less in 
this case, so that mixing ought to be much less. The data of George Oliver 
will be shown to prove that the mixing, in fact, is much less for a 
nonshrouded valve than for a shrouded valve. 
Other data from Lancaster's paper strongly reinforces the idea of the flow 
inside the cylinder of an engine during the intake and compression stroke 
as a hydrodynamic "dance" with random turbulence superimposed on the very 
nonrandom mean flow pattern. FIG. 20 (Lancaster's FIG. 6) shows effects of 
engine speed on ensemble averaged mean velocity with the shrouded intake 
valve. The similarity of the flow structure (hydrodynamic dance) from 
engine speed to engine speed is striking. FIG. 21 (Lancaster's FIG. 7) 
shows the effect of engine speed on the ensemble averaged mean velocity 
for the nonshrouded intake valve case. Again, the similarity of the 
hydrodynamic dance from speed to speed is striking. 
FIG. 22 (Lancaster's FIG. 8) shows the effect of changing volumetric 
efficiency on the ensemble averaged mean velocity with a nonshrouded 
intake valve for a set engine speed. The shape of the mean velocity is 
much the same from one throttle setting to another. FIG. 23 (Lancaster's 
FIG. 9) shows the effect of volumetric efficiency changes on ensemble 
averaged mean velocity for the shrouded valve for a set engine speed. 
Again, the concept of a basically reproducible and stable hydrodynamic 
dance is very strongly supported. 
A comparison of FIG. 18c and FIG. 19 shows clearly that the velocity 
gradients of the flow structure in the cylinder are significantly greater 
with the shrouded than with the unshrouded valve case. Consequently, 
mixing of the shrouded valve engine must be much better than that with the 
unshrouded valve case. The work of George Oliver very strongly shows this 
mixing difference, and gives vivid demonstration of just how heterogeneous 
the charge inside an engine with a nonshrouded valve can be (Cloud 
Combustion, A Study of Performance and Emissions, M. A. Theses, Chemican 
Engineering Department, California Institute of Technology, June 4, 1973). 
Oliver's work was done on an engine essentially identical to the engine 
for which Lancaster took his flow structure and turbulence measurements. 
Oliver ran an engine where the mixture in the intake passage was 
stratified by means of gas injection into a long intake tube. Since on a 
one-cylinder engine, flow in the intake passage occurs only about one 
quarter of the time, injecting the fuel (propane) continuously at a 
specific length away from the intake in the intake tube assured that most 
of the fuel would enter the cylinder as a plug (with fuel smeared out due 
to the fact that a quarter of the fuel was injected with the gas column 
moving, and also smeared out because of viscosity effects). Running this 
engine with a shrouded valve showed that the combustion performance of the 
engine was insensitive to this intake stratification. However, with a 
nonshrouded valve, Oliver's data shows an extremely marked combustion 
effect due to stratification in the intake passage. Oliver's data strongly 
shows that mixing in the cylinder was sufficiently slow that a definite 
charge stratification in the cylinder existed at ignition time because of 
intake flow stratification. Oliver's data is, therefore, evidence both for 
definite and reproducible flow structures in the engine and for the 
relatively very slow mixing characteristic of nonshrouded valve 
conventional intake geometries. For basic physical reasons (because the 
mixing events in-cylinder are dominated by intake induced flows) there is 
strong reason to expect that mixing in conventional engines as they are 
currently built is much like that in Oliver's engine. For this reason, 
conventional engines require much richer mixtures than would otherwise be 
necessary (for reasons explained graphically in FIG. 6). 
A high swirl flow inside an engine does not necessarily mix well, as the 
description of FIG. 15 points out. It should be pointed out that flows 
from swirl ports quite frequently do produce a flow which is very close to 
rigid body charge rotation. Data showing this beyond question was 
presented in "Measurement of Air Movements in Internal Combustion Engine 
Cylinders," by M. Horvatin and A. W. Hussmann in DISA INFORMATION, #8, 
July 1969. FIGS. 24 and 25 are Horvatin and Hussmann's FIG. 5. FIG. 24 
shows the cup in piston and cylinder design of the diesel engine tested. 
FIG. 25 shows positions of hot wire aenemometer sensors in the cylinder 
head, and shows shadow lines showing the shape of the intake and exhaust 
ports. FIG. 27 is Horvatin and Hussmann's FIG. 7, showing the measured 
flow pattern in the cylinder head 10 mm. below the cylinder head. The FIG. 
27 contains its own explanation: note particularly how closely the flow 
pattern during the intake and compression stroke resembles rigid body 
rotation. This flow pattern which approximates rigid body rotation is 
characteristic of a number of swirl ports, and does well with a diesel 
engine. However, it is very bad for mixing during the intake and 
compression stroke. 
It is expected that the FIGS. 3-27 have explained that mixing and 
structured flow are important in engines, shown how different flow 
structutres have very different mixing properties inside combustion 
chambers, and shown that different engines have flow structures having 
very different mixing properties. These are points which have not been 
clearly understood by the automotive engineering fraternity, and points 
which are most important to the function of the present invention. 
STRUCTURED VORTEX FLOW WITH PRESENT INVENTION ENGINE 
At this point, a competent automotive engineer should be able to appreciate 
the strong evidence for an irrotational flow vortex much like that shown 
in FIG. 10 in the inventor's test engine, shown as a photographic 
reproduction as FIG. 28. This photograph shows the deposit pattern on the 
flat top piston used with the engine for which much data will be included 
in this application. The deposits were caused by an oil control problem 
not related to the present invention. The same deposit pattern occurred on 
several engine teardowns. The very clear deposit pattern on the piston 
top, as a result of some hours of engine operation, shows vividly that 
there was a structured flow inside the combustion chamber. A knowledge of 
fluid mechanics permits the pattern to be clearly interpreted: the 
accumulated deposit at 13 and the spiral pattern in toward this center is 
characteristic of the boundary layer flow of a surface perpendicular to 
the axis of an irrotational flow vortex. This boundary layer flow occurs 
because of a depressed pressure in the center of an irrotational flow 
vortex. Within the boundary layer, velocities are sufficiently low that 
centrifugal force is relatively much less significant than outside the 
boundary layer, and consequently, oil in the boundary layer is sucked into 
the low pressure center of the vortex, where it evaporates. A conventional 
rigid body rotation vortex has a much weaker pressure depression in its 
center, and would not produce this pattern. The deposit markings of FIG. 
28 are conclusive evidence of a strong irrotational velocity component in 
the flow structure inside the engine's combustion chamber. In light of the 
mixing structure produced in irrotational flow vortices and illustrated in 
FIGS. 10, 11, 12, 13, and 14, it is clear that mixing, at least within the 
volume of the large vortex inside the engine's cylinder volume, must be 
very complete. Other evidence substantiates that the mixing inside this 
engine of the present invention was indeed excellent. 
EVEN WITH PERFECT MIXING, TURBULENCE IS REQUIRED AT IGNITION TIME FOR FLAME 
SPEED 
Excellent mixing is required to consistently burn the very dilute mixtures 
required for ultralow NO.sub.x emissions without catalytic reduction of 
NO.sub.x. However, for the resulting combustion to be commercially useful, 
combustion duration must be fast enough. With conventional production 
engines, the dilute mixtures required for low NO.sub.x burn so slowly that 
the thermodynamic losses due to late heat addition overshadow the 
theoretical cycle advantages of dilute combustion. This is a large part of 
the reason why the idea that there is a fuel consumption penalty 
associated with dilute combustion has become entrenched. However, the 
speed at which a mixture is burned in an engine is not a simple function 
of concentration of fuel, air, and residual gas. In conventional engines, 
flame speed for a set air-fuel ratio varies by about a factor of eight for 
a ten to one variation in engine speed. The reason is that flame speed 
depends on turbulence. 
The best work on this flame speed versus turbulence relation is probably 
"Effects of Turbulence on Spark-Ignition Engine Combustion" by David R. 
Lancaster, Roger B. Krieger, Spencer C. Sorenson and William L. Hull, SAE 
paper 760160. In this paper, the flow and turbulence measurements of 
Lancaster's SAE paper 760159 are applied to flame speed. The most 
important result of this paper is reproduced as FIG. 29 (SAE paper 
760160's FIG. 15) which plots the flame speed ratio Turbulent flame 
speed/Laminar flame speed versus the root mean squared turbulence 
intensity. It is important to point out that this turbulence intensity is 
the root mean square value of the random fluctuations from the flow mean 
velocity, and would correspond to the "intensity" line of FIG. 18c. 
FIG. 30 (FIG. 17 of SAE 760160) shows the correlation Lancaster and 
associates found between the flame speed ratio (of turbulent to laminar 
flame speed) and the inlet flow velocity (which was proportional to RPM). 
The conclusion of Lancaster et al. vis-a-vis FIG. 30 was that for a set 
engine geometry, flame speed correlated with inlet velocity. 
The correlation between flame speed and inlet flow velocity means that in 
order to increase the flame speed ratio to compensate for the slower 
laminar flame speed of very dilute mixtures, the area through which the 
inlet flow comes into the combustion chamber must be decreased. For a 
fixed cam timing conventional engine, this requires that peak engine flow 
capacity be restricted, but of course this power-flame speed trade-off 
goes away if the inlet flow restriction is variable. 
TIGHT STATISTICS AND CONTROLLED STRUCTURED FLOW MAKE EFFICIENT LOW NO.sub.x 
ENGINE POSSIBLE 
The matters which have been discussed concerning NO formation, mixing, and 
flame speed are complex and interrelated. The conceptual complexity 
becomes greater when the problem of producing mass producible engine 
hardware is fully understood. It is therefore not surprising that the 
automotive industry has decided (and frequently testified under oath) that 
very low NO operation with a homogeneous charge engine is impractical 
without catalytic reduction of NO, or at least substantial performance 
penalties. 
However, it is in fact possible to achieve ultralow NO.sub.x outputs and 
excellent performance with a homogeneous charge engine format. This is 
accomplished according to the present invention by producing an engine 
which combines the following: (1) a very dilute fuel-air mixture, (2) very 
tight cycle-cycle and cylinder-cylinder mixture statistics to each 
cylinder, (3) a variable and highly structured intake flow pattern which 
produces sufficient turbulence for rapid flame propagation with very 
dilute mixtures. An engine with these attributes in combination will have 
both extremely low NO emissions and excellent fuel economy, as data in 
this application will show. With the intake flow controlling restriction a 
variable restriction, and with full power mixture enrichment, this very 
low emission operation is completely compatible with outstanding peak 
engine power for the relatively rare times when peak engine power is 
required of an automobile engine. Also, the controlled flow structure and 
turbulence permits the low NO.sub.x operation to be achieved with 
excellent engine smoothness. 
It is important to emphasize that the structured flow patterns's mixing 
function and the structured flow's combustion turbulence generating 
function are distinct. The mixing statistics and the detailed mixing 
stratification function in space inside the cylinder are the result of the 
history of the flow, including all the details of mean flows and 
turbulence, up to the instant in time where the mixture state is 
considered. Mixing is the effect of turbulence and flow structure 
integrated over time. The turbulence which determines the turbulent to 
laminar flame speed ratio (and therefore the actual flame speed with a set 
ratio of fuel, air, and residuals) is the turbulence actually present at 
combustion time. 
The detailed fluid mechanics during the entire intake and compression 
stroke is important for mixing. Only the fluid patterns actually existing 
during combustion time are important in determining turbulent to laminar 
flame speed ratios. This is an important distinction in comparing the 
present invention with other means of increasing flame speed (for example, 
air jets from auxiliary pistons and squish pistons) which may have large 
effects of flame speed and yet operate with much less tolerance for dilute 
mixtures because of inferior mixing. 
DETAILS OF FLOW GEOMETRY CRITICAL FOR MIXING, TURBULENCE, AND STABILITY 
Another issue with respect to turbulence is of very great importance. The 
rate at which flow kinetic energy decays to heat (internal energy) varies 
drastically from flow geometry to flow geometry. For properly designed 
intake port passage shapes, flow patterns are such that the flow energy 
past a variable port restriction organizes itself in a fashion such that 
differential velocities between adjacent fluid elements are minimized, 
with smooth velocity gradient variations. Under these conditions, 
significant fractions of the pumping energy flowing past a variable 
restriction can be organized in such a way as to generate flow patterns in 
the cylinder which produce very excellent mixing, and the flow energy in 
these patterns can decay into turbulence slowly enough so that large 
amounts of turbulence are available to facilitate combustion at and after 
ignition time. A number of such shapes will be disclosed in the present 
application. However, such flow structuring is not particularly likely to 
happen by accident: many flow geometries involve such rapid decay of flow 
kinetic energy to fine grain turbulence that they are effectively useless 
for mixing and increasing flame speed at combustion time, since the 
vortices are too small scale for large scale mixing, and decay too rapidly 
to assist combustion. Other flow geometries are only marginally useful. 
The inventor has spent a considerable amount of time with hot wire 
aenemometry steady flow setups investigating flow in a cylindrical passage 
downstream of an engine cylinder head, investigating various variable 
restriction inlet flow geometries. The inventor was most surprised to find 
that the flow energy from a number of variable restriction inlet 
geometries decayed so fast that, with these geometries operated with 
significant restrictions, turbulence in the cylindrical channel about 10 
cm. downstream of the head was less intense than the turbulence which 
occured without the restriction using a conventional nonshrouded intake 
valve (these tests were conducted for the most part on a modified 350 CID 
Chevrolet cylinder head). Unless the flow geometry is correct, a variable 
restriction intake port can be useless or worse than useless. 
In addition, there are inlet flow geometries which produce useful results 
if shapes are exactly correct, but which produce bad results with only 
very small changes in passage shaping, so that they are impractical to 
manufacture. Moreover, even for basically good geometries, certain shaping 
issues are critical if the variable restriction intake port passage is to 
operate properly. 
To think about the fluid mechanical effects relevant to controlled 
structured flow via variable restriction intake port geometries an 
automotive engineer will have to turn to studies foreign to conventional 
automotive engineering, and particularly to the study of fluidics. 
NECESSARY FLUIDICS BACKGROUND 
Fluidics is a field which applies fluid mechanics to information handling. 
The development of fluidics has greatly advanced the development of the 
conceptual tools required to think about structured turbulent flows. The 
inventor was primarily instructed in fluidics by Raymond Warren, one of 
the founders of the field. The conventional fluid mechanical training 
taught to engineers generally treats turbulent flows as predominantly 
random: in fluidics it is shown that, if the passage shaping relationships 
are right, reproducible flow patterns such as wall attached streams and 
stabilized vortices can dominate the flow, with the smaller order 
turbulence of much secondary importance. Applying the fluid mechanical 
effects used in fluidics to intake ports is not a trivial exercise: 
problems arise due to the manufacturing variations to be expected of 
intake ports, and due to the fact that an intake port is a 
three-dimensional flow passage much constrained in shape both at its inlet 
and outlet, and a flow shape where flow restriction of the passage is 
important. Moreover, essentially all fluidic devices have fixed geometry, 
and the variable port restriction geometry varies. However, an 
understanding of fluidic effects is vital to the function of the present 
invention, and this understanding can best be attained by considering 
practical fluidic devices. 
FIGS. 26a, 26b, 26c, and 26d (taken from FIG. 1.16, Page 18, FLUIDICS, 
COMPONENTS AND CIRCUITS, by K. A. Foster and G. A. Parker) show a number 
of important fluidic effects in a compact way, by showing the function of 
Parker and Jones' cusp half adder binary fluidic device. FIG. 26a shows 
the device with its flow pattern in the A=0, B=1 state; FIG. 26b shows the 
device with its flow pattern in the A=1, B=0 state; and FIG. 26c shows the 
device in its A=1, B=1 (carry) state. FIG. 26d shows the passage shape of 
the half adder without flow streamlines to emphasize the critical 
dimensions relevant to this device. 
An observer must be struck, in observing FIGS. 26a-c, by the extremely 
structured flow pattern produced by this flow geometry, with the momentum 
of the streams effecting the flow, and with vortices stabilizing in cusps 
in the flow passages, drawing flow energy from the main stream(s) in the 
manner of a roller engaging a surface, and with the parasitic vortices 
actually serving to stabilize and organize the flow. It should be 
emphasized that there is turbulence everywhere in the fluid, so that the 
flow lines are mean flow lines. Nevertheless, the mean flow lines show the 
much predominant flow patterns, on which turbulence superimposes as small 
perturbations of relatively little importance. The flow has turbulence, 
but it is very far from random. 
Virtually, all of the fluid mechanics shown with repect to Parker and 
Jones' device of FIGS. 26a-26d is relevant to the function of the fluidic 
ports which enable continuously variable flow and turbulence to be 
produced in an engine. Effects which must be kept in mind are the 
following: 
a. As the fluid elements move, they follow a path such that the sum of the 
forces acting on each element (including inertial forces) is balanced, in 
a way well known to designers of turbomachinery. In rectangular 
coordinates: .epsilon.F.sub.x =0, .epsilon.F.sub.y =0, .epsilon.F.sub.z 
=0. Much of the shaping of port passages for efficient fluidic performance 
depends on this simple dynamics. 
b. Turbulent streams interact with adjacent fluid, and tend to entrain 
fluid and thereby spread, with reductions in velocity and increases in 
stream mass (and with, as a first approximation, conservation of 
momentum). When a high velocity jet is near a surface, the surface blocks 
flow into the jet and, therefore, the pressure on the wall side of the jet 
is reduced. The consequent pressure imbalance across the jet forces the 
jet toward the wall (at a rate so that inertial forces balance the 
pressure imbalance): the jet is "sucked" toward the wall. As the jet comes 
closer to the wall, the pressure difference forcing the jet toward the 
wall increases, so that the jet tends to lock onto the wall and becomes a 
wall attached stream (Coanda wall attachment effect). 
When the stream is attached to the wall, the jet can only entrain fluid on 
its free side, and so the jet entrains less fluid and consequently spreads 
more slowly. Over a unit of flow distance, the reduction in spreading is 
more important than the viscous drag of the surface, so that a wall 
attached stream conserves its kinetic energy with distance downstream much 
better than an unattached jet. Also, if the high speed jet is shaped so as 
to minimize its free side surface area, entrainment effects will be 
minimized and energy conservation as the flow moves downstream will be 
better. 
c. As was illustrated in FIGS. 26a, 26b, and 26c, a high velocity stream 
can induce, in a properly shaped flow passage, rather stable vortices or 
zones of recirculation, which are driven by the main flow. These 
recirculation patterns can very much stabilize the action of a wall 
attached stream, and by greatly reducing the velocity gradient between the 
wall attached stream and other flow in the port, the secondary 
recirculation flow can greatly decrease entrainment of fluid into the high 
speed jet, and thereby much reduce spreading of the jet. By reducing the 
spreading of the jet, the fraction of maximum jet velocity and kinetic 
energy which can be recovered in coherent form can be much increased over 
the fraction that would obtain without these zones of recirculation. For 
this reason, portions of the flow passage well away from the attached 
stream can change the spreading angle of a wall attached jet in the flow 
passage, by changing the flow and stability of the recirculation pattern. 
d. The stability of a wall attached stream is a function not only of the 
intrinsic instabilities which come from high local Reynolds numbers: small 
variations in flow geometry which might at first appear to be 
insignificant can, by producing large disturbances in the jet, cause a 
magnification of turbulence causing the jet to "break up" in a way causing 
great reductions in the fraction of the velocity and kinetic energy of the 
jet as it travels downstream. In FIG. 26d, dimensions important to this 
stability are shown both as cusp setback y and wall set back x. The issue 
of setback is important to the successful and reliable function of 
fluidically efficient variable restriction ports. Streams which flow past 
a setback will, unless the setback is too great, reattach cleanly to the 
wall. An example of this is shown in FIG. 27a. Near the setback, as shown 
in FIG. 27a will be a recirculation zone or vortex 30 known in the 
fluidic arts as a "separation bubble." In FIG. 27a the setback dimension a 
is shown. As setback is decreased, the separation bubble shortens and the 
fraction of the fluid energy given up to the separation bubble decreases. 
However, if the setback becomes a "step-up," a high velocity element of 
fluid will collide with the stepup as with a brick wall, and will deflect 
in a way which strongly perturbs the flow, breaking up and destroying the 
coherence of the jet, and so ruining the fluidic efficiency of the device. 
See FIG. 27b. The high velocity jet 32 hits the stepup 33 and breaks into 
intense turbulence, spreading at such an angle that the pressure recovery 
from the jet downstream will be very small. Once a jet is well attached to 
a surface, surface roughness (for instance sand casting roughness) has 
surprisingly little effect. However, a stepup, or an intervening gasket 
protruding into the channel at the jucture between two passages, can 
completely break up the fluidic pattern which would otherwise occur. In 
the present invention, this stepup effect is important in two different 
ways. For fluidic efficiency of a restriction-caused jet into the 
combustion chamber, stepups must be avoided, and for mass production this 
means that parts must be assembled with tolerances which always involve 
setbacks and never involve stepups on the passages relevant to flow into 
the combustion chamber. Also, to eliminate perverse wall attached jets of 
backflow into the intake manifold during the valve overlap period, a 
stepup with respect to backflow near the variable restriction is effective 
in destroying the coherence of the backflow jet, and so simplifies intake 
manifold design. 
A point of great importance with respect to fluidic principles applied to 
intake ports must be made. The inventor had great trouble with this 
initial port designs, which initially worked very well and then with what 
seemed to be very small geometrical changes did not work at all (or worse, 
some cylinders worked well and others worked badly). Detailed flow 
patterns, on which mixing and flame speed in the engine depends, can be 
metastable or bistable in the port unless geometries are right. This is 
often a manner of details, and a very large part of the inventor's effort 
since the filing of this application's parent case has been development of 
fluidic port designs that are stable, so that small variations in passage 
geometry did not have catastrophic effects on the function of the ports. 
Many fluidic devices are designed to be metastable, since they often 
function as binary switches. Even so, the sensitivity of the flow to 
geometry can be very great, and not easily appreciated by one skilled only 
in the conventional automotive engineering arts. A quote with respect to 
sensitivity of a binary fluidic switch should emphasize the point: 
"The problem, however, was that at small setback distances, the sensitivity 
changed fairly rapidly so that the effect of errors in the manufacture of 
an element could create great differences in characteristics; this is 
particularly noticeable in that similar elements from a batch may vary 
from monostability on one side through bistability to monostability on the 
other side, even though they appear superficially to be accurately made." 
(Page 314, Foster and Parker, op.cit.). 
Such sensitivity is clearly unacceptable in a mass produced automobile 
part, and is particularly unacceptable in a sand cast geometry such as an 
intake port. Therefore, the design of the intake port shape for stability 
is a vital issue. In general, to achieve stability, all the fluidic 
effects must be such that they strongly push the flow into the desired 
pattern, so that the flow is insensitive to small variations in geometry. 
This can be done so long as stepup is avoided: in the inventor's 
experience, stepup, unless it was vanishingly small, has always been 
disastrous to the fluidic efficiency of a port geometry. If the basic 
shape is correct and stepup is avoided, it has been found that fluidically 
efficient ports can be made with sand cast surface quality and within the 
range of geometrical variations to be expected in mass production 
castings. 
FLUIDICALLY STABLE VARIABLE PORT RESTRICTION GEOMETRIES 
See FIG. 31, which is a side view showing an inwardly opening flap 
restriction which opens on the floor of an intake port. Flap 35 hinges on 
pivot shaft 36, and swings inwardly inside port 37, sealing along the 
sides of the flap (not shown) and forming an opening between the bottom of 
said flap 35 and port floor 38. Flow toward the restriction has vector 
components onward along the y axis as well as along the x axis, and in 
consequence, the flow contracts past the restriction according to the well 
known vena contracta effect to form a minimum flow cross section less than 
the restriction cross section at vena contract point 39. The jet flow at 
39 has a velocity very closely corresponding to the isentropic velocity 
caused by the pressure drop across flap 35. For straightforward inertial 
reasons, the flow jet attaches to the floor of the port 38: no switching 
time is required for this attachment process, since inertia forces the 
stream against surface 38. The high velocity jet 40 will entrain flow, and 
impart momentum to the air inside the port passage, which will rapidly 
form a recirculating vortex 41: this recirculating vortex will strongly 
increase the fluidic efficiency of the port by reducing the velocity 
gradient between the top of jet 40 and the remaining fluid inside the 
port, and so reducing the spreading angle of jet 40 and increasing the 
velocity which will be delivered across the valve seat 42. The jet 40 
detaches relatively cleanly at curvatures 43, and rushes across the valve, 
so that the great majority of the flow happens across only a small 
fraction of the valve opening area, in a way which produces structured 
flow and controlled mixing inside the combustion chamber of the engine 
(not shown) and, therefore, permits stable and rapid combustion with the 
very dilute mixtures required for NO.sub.x control and improved fuel 
economy. Clearly, by pivoting flap 35, the flow energy delivered into the 
engine combustion chamber (not shown) can be varied over a wide range, and 
when the flap is fully open, very low flow resistance is possible with 
good passage design. 
The flow pattern of FIG. 31 illustrates important points about the use of 
fluid momentum to assure that a high speed jet attaches to the desired 
surface according to the Coanda effect. FIG. 31 also shows the parasitic 
vortex patterns inside the port geometry which reduce spreading, and so 
increase port fluidic efficiency. However, FIG. 31 is somewhat 
oversimplified, because it takes a fundamentally two dimensional view of a 
fundamentally three-dimensional flow passage. Furthermore, jet 40 jumps a 
gap as it goes across the valve opening, and this jump involves flow 
passage shaping issues which are geometrically touchy. Because of the 
geometrically touchy nature of curvature 43, it is likely that ports 
having a flow jump across the valve opening, such as that which occurs 
from curvature 43, will be difficult to mass produce. 
For reasons concerning the geometry of a swinging door type flap 
restriction, fluidic ports work best when they are of generally 
rectangular cross section. A high speed stream attached to one wall in a 
rectangular passage will tend to move into one of the two corners adjacent 
to the wall to which it is attached. For straightforward fluidic reasons, 
the stream would rather be attached to two adjacent walls in a compact 
pattern than in a less compact pattern spread out over one flat wall. This 
preference of the stream for corners can be very useful in the design of 
fluidic ports. However, the issue of which corner of the rectangular 
passage the stream flows to is vital, and if the flow pattern is 
metastable, with the flow sometimes attaching to one wall and sometimes 
attaching to another, the effect can be disastrous. 
FIG. 32 is a perspective cutaway view of a rectangular passage with a 
bottom opening flap, (flap not shown) so that the high velocity flow will 
attach to the floor of the passage 45. Section lines a, b, and c show 
cross section cuts which are shown in FIGS. 33a, 33b, and 33c. In FIGS. 
33a-c, the shaded area is the area containing the high speed wall attached 
flow: clearly, the shapes and edges of this high speed flow are somewhat 
arbitrary, since there will be smooth velocity transitions within the 
flow--the edges of the shaded regions could, for example, indicate the 
flow at 60% of the vena contracta velocity. However, the major point of 
the FIGS. 32 and 33a-c is that the wall attached stream tends to contract 
itself into a minimum surface area configuration and go into a corner of 
the passage. 
FIG. 34a shows actual measured results with a port arranged as shown in 
FIGS. 34a, 34b, and 34c to flow in the manner explained in FIGS. 32 and 
33a-c. FIG. 34a is a view of the combustion chamber area of the head with 
the view perpendicular to the plane of the head cylinder sealing surface 
and showing the general shape of the intake port passage 50 in shadow 
lines. The numbers and vectors shown around the intake valve show actual 
measured percentages of the isentropic velocity past the flow resctiction 
54 in this particular setup for steady flow tests. FIGS. 34b and 34c are 
additional views of the intake port passage and variable restriction. FIG. 
34b is a cutaway of the intake port generally along the center line of the 
flow of the port and shows rectangular throttle 54 in relation to port 
section 50 and adjoining intake manifold passage 52. FIG. 34c is a cutaway 
view of the same head arrangement on a cutaway plane passing through the 
center lines of the intake and exhaust valve stem. The port arrangement 
shown was modified from a 350 cu. in. displacement Chevrolet engine head. 
Velocity measurements were taken with Pitot tubes modified from medical 
syringe needles. Pressure differences were measured with water inclined 
manometers. 
For efficient fluidic ports designed so that the flow goes into a corner of 
the generally rectangular port passage, it is most important that the flow 
always, even with small production variations from passage to passage, 
goes into the proper corner. Unless this is done, the design cannot be 
commercial. 
The following drawings show ways to insure that the flow goes into the 
proper corner so that it is delivered in coherent form into the cylinder. 
FIGS. 35, 36a, 36b, 37 and 38 are each drawings for a system having 
outwardly swinging flaps and where the outwardly swinging flaps are part 
of a separate intake manifold assembly which mounts to the intake port on 
the cylinder head in the manner of FIG. 39 which will be discussed in more 
detail subsequently. FIG. 35 shows how the flow can be biased to attach to 
a specific corner by the simple expedient of arranging the setback so that 
the setback on one wall is significantly less than the setback on the 
other. FIG. 35 is a view looking in the direction of the port runner of 
the intake manifold near where it mounts to the cylinder head for a 
specific port and showing the swinging flap 60 in the closed position 
generally perpendicular to the runner. Dotted line 61 denotes the outline 
of the intake port in the cylinder head against which the manifold section 
would mount. The alignment of the manifold passage and the intake port is 
such that there will be setback all around the intake port. However, the 
setback B on the left side of the port will be significantly less than the 
setback A on the right side of the port. In consequence, the flow from the 
bottom opening flap will tend to attach to the port runner on the lefthand 
corner port side. 
FIGS. 36a and 36b illustrate the use of deflectors to deflect the wall 
attached stream from the variable restriction towards the desired intake 
port corner. Manifold runner passage 70 mounts variable restriction flap 
72, which pivots on control shaft 73. The manifold passage 70 mounts onto 
an intake port runner 74, which is integral to an engine cylinder head not 
shown. On the floor of intake manifold passage runner 70 are deflector 
fins 76 which serve to deflect the flow towards wall 79 so that it will 
attach on the corner between the floor of the port and wall 79. Again, 
note that the juncture between manifold passage 70 and port runner 74 has 
setback. 
FIG. 37 shows another very simple way of biasing the flow toward a specific 
corner by arranging the predominant fraction of flow when the flap 60a is 
relatively closed on the left side by providing an opening along the left 
side of a variable restriction flap section 80. 
FIG. 38 illustrates how the wall attached stream can be biased to the 
proper wall by means of a deflection stream. Variable restriction flap 86 
includes a flow passage 87 on the downstream face of the flap 86 which 
connects flow from the upstream side of the flap through hole 88 to be 
delivered in the proper deflecting direction at opening 89. 
Those skilled in the fluidic arts will recognize that there will be other 
ways to assure that the flow stream attaches to the proper wall of the 
port runner and that FIGS. 36 to 38 are exemplary only. 
It has been discussed before that the variable port restrictions serve to 
control the blowback of exhaust gas during the intake and exhaust valve 
overlap period, as well as control the fluid mechanical details of the 
intake flow into the cylinder. Referring specifically to FIGS. 36a and 
36b, the opening of the variable restriction flap 72 shown by height H in 
FIG. 36a, is not only the flow cross section through which all intake flow 
must pass; gap H is also the flow path through which any exhaust gas from 
the intake port must flow in order to flow into the intake manifold port 
runners. Therefore, it is very clear that changing the angle of flap 72 to 
reduce gap H will reduce the internal exhaust gas recirculation for a 
given engine speed and load which is produced with a set valve overlap. 
Therefore, controlling the angle or opening of the variable restrictions 
at each intake port will serve as a programmable internal EGR control 
device. Because the port restrictions can be very restrictive, the 
variable restrictions will permit engines to be built with very 
substantial intake and exhaust valve overlap periods, since the mass flow 
of exhaust gas recirculation at low speeds and loads can be completely 
controlled by the variable restriction settings. The ability of the 
variable restrictions to permit very high valve overlaps eliminates one of 
the most important trade-offs between engine smoothness and peak engine 
power in prior art engines. In addition, the variable restriction in 
conjunction with high overlap camshafts (which are also desirable for peak 
power) will provide an extremely efffective programmed EGR device and will 
eliminate the need for any external EGR plumbing. The elimination of 
external EGR valves and plumbing will represent a substantial cost saving 
to engine manufacturers. However, it has been found experimentally that 
the exhaust blowback flow during the valve overlap period can produce very 
bad cylinder-to-cylinder distribution of residual gases unless the flow 
downstream of the variable restriction during the valve overlap period 
(which during the intake stroke is upstream of the valve restriction) is 
properly arranged to break up Coanda wall attached streams. The function 
of stepup in breaking up a Coanda wall attached stream has already been 
discussed. It has been found experimentally that provision of a stepup 
such as 78 (which is a setback during the intake flow) will effectively 
break up the blowback flow so that it dissipates into turbulence and does 
not penetrate too far into the intake manifold assembly. The provision of 
this step 78 is very important for the practical operation of the variable 
restriction system with significant valve overlap. Without the breakup of 
the exhaust blowback wall attached stream, it is very much more difficult 
to build an intake manifold arrangement having adequate 
cylinder-to-cylinder exhaust gas distribution and proper fluid mechanics. 
It should be noted that stepup 78 is very different in function from the 
steps sometimes provided in intake manifold assemblies to reflect pressure 
waves. A wall attached stream is clearly not a pressure wave. 
For manufacturing reasons, it will be much more convenient to have the 
variable restriction arrangement part of the intake manifold assembly 
rather than part of the cylinder head. Cylinder head manufacture is 
conventionally done by sand casting, which does not lend itself to super 
close tolerances or to smooth surfaces. The intake manifold assembly, 
however, can be arranged as an assembly of die case aluminum parts which 
can be made inexpensively to much higher tolerances than are obtainable 
with the cylinder head. FIG. 39 illustrates these points. A cylinder head 
100 has attached to it an intake manifold assembly 102-110 including the 
variable restrictions. Manifold main section 102 includes within itself 
intake manifold passage runners 103 corresponding to each intake port and 
any step ups or fins which are fluidically worthwhile (such as those 
illustrated in FIGS. 35a and 35b). The manifold main section 102 has a top 
section (also a die casting) 106 and at the parting line between 102 and 
106 is accomodation for a flat shaft assembly 104 which mounts the 
variable restriction flaps 105 for each intake port. A homogeneously mixed 
vaporized fuel air mixture (produced in a way which will be discussed 
later) is delivered to the intake manifold through inlet section 108. The 
control of the flap shaft angles as a function of carburetor throttle 
angle and other engine parameters is controlled by control means 110 which 
is not shown here in any detail. Die cast manifold assemblies such as 
those shown in FIG. 39 may be adapted to both straight and V engine types, 
and significantly increased geometrical complexity is possible with the 
die cast technique. However, the realities of engine manufacture make it 
very desirable to arrange the variable restriction arrangement as an 
integral part of the intake manifold assembly so that the entire variable 
restriction arrangement (excepting simple core changes in the intake port 
shape) is part of one assembly which may be put together on a sub assembly 
line in a manner which causes minimum disruption of the engine assembly 
line itself. 
The variable restriction flap arrangement of FIG. 39 is adapted for 
outwardly swinging flaps such as those shown in FIGS. 36 and 36b. However, 
other variable restriction arrangements are possible. Two alternative 
arrangements are illustrated in FIGS. 40 and 41. FIG. 40 illustrates a 
variable restriction arrangement having a sliding vane 120 to control the 
flow. FIG. 41 illustrates a variable restriction flap which swings in 
toward the intake port. An inwardly swinging flap has small advantages 
with respect to the fraction of the isotropic velocity past the 
restriction delivered in coherent form past the intake valve face. 
However, inwardly swinging flaps in manifold assemblies are less compact 
than outwardly swinging flaps. 
The variable restriction intake ports do an excellent job of homogeneously 
mixing the fuel, air and residual gases in the cylinder. However, no 
amount of in-cylinder mixing can correct for variations in the air fuel 
ratio or fresh charge residual ratio delivered to the cylinder from cycle 
to cycle, nor can incylinder mixing correct for variations in air fuel 
delivery or residual charge delivery from cylinder to cylinder. Those 
skilled in the art of combustion engines will recognize that getting 
completely perfect cylinder to cylinder mixture distribution is a very 
difficult thing. The statistical argument in FIG. 6 was intended with 
respect to microscale mixing in the cylinder. However, this argument is 
just as valid with respect to cylinder to cylinder or cycle to cycle 
statistical variations. To take full advantage of the fluidic variable 
restriction ports (that is to say to go very lean for maximum nitric oxide 
reduction) it is necessary to have homogeneously mixed charge delivered to 
each cylinder. 
Apparatus for producing very tight in-cylinder species concentration 
statistics and for increasing flame speeds in engines have now been 
disclosed. However, with respect to the critical issue of nitric oxide 
formation, the chemical events governing nitric oxide formation require 
that the species concentrations of fuel, air, and residual gas at each 
point in the combustion chamber for each cylinder be dilute enough to 
produce the low NO.sub.x levels. No matter how good the in-cylinder mixing 
is, relatively high NO.sub.x levels and/or relatively high misfire levels 
can occur if too much statistical variation in fuel-air-residual delivery 
occurs between cylinders or cycle to cycle for a given cylinder. For 
efficient, low NO.sub.x, dilute combustion, very tight 
cylinder-to-cylinder and cycle-to-cycle mixing statistics are required in 
addition to the requirement for in-cylinder mixing. The overall cylinder 
air-fuel-residual gas ratio is determined at the end of the intake stroke, 
and no amount of in-cylinder mixing can change this ratio. Therefore, the 
requirement for tight cylinder-to-cylinder and cycle mixture delivery 
statistics is inescapable. As has already been discussed, the requirement 
for microscale in-cylinder mixing is also inescapable if dilute combustion 
is to be achieved. 
Conventional engines have relatively bad cylinder-to-cylinder and 
cycle-to-cycle fuel-air delivery statistics, particularly under transient 
operation conditions and under cold start conditions. Referring back to 
the statistical illustration of FIG. 6 and considering the source of 
statistical deviation to be cycle-to-cycle or cylinder-to-cylinder 
variations, it is clear that the tighter the cylinder-to-cylinder and 
cycle-to-cycle mixing statistics the more dilute combustion mixtures can 
be without unstable combustion or misfire, and therefore the lower 
NO.sub.x emissions can be and the better engine fuel efficiency can be. 
It is possible, although technically quite difficult, to get very tight 
cylinder-to-cylinder and cycle-to-cycle mixing statistics with fuel 
injection nozzles at each intake port. Even with perfect 
cylinder-to-cylinder distribution at the injection nozzles, much of the 
liquid fuel wipes out on port surfaces and there is a suprisingly and 
inconveniently large lag under transient conditions due to the two-phase 
flow in the port sections, so that even the best port fuel injection 
systems are less than perfect with respect to delivery characteristics to 
the cylinder under transient conditions. In addition, individual port fuel 
injection is quite expensive and tends to be difficult to maintain. For 
these reasons, the inventor, working with co-workers Charles L. Siewert 
and Kenneth W. Kriesel, has developed a vortex fuel-air mixing device 
employing turbulent structured flow fluid mechanics to produce the 
required steady state and transient fuel-air distribution statistics. This 
vortex mixer is the subject of a co-pending patent application, Ser. No. 
945,388 now abandoned in favor of continuation-in-part application Ser. 
No. 77,759 filed July 20, 1979. However, means to economically achieve the 
very tight cylinder-to-cylinder and cycle-to-cycle statistics required for 
very low NO.sub.x emissions are sufficiently important and the structured 
flow fluid mechanics effects used by the vortex mixer are sufficiently 
important, that the device is discussed in detail here. 
FIGS. 50, 51 and 52 show a preferred form of the vortex mixer. FIG. 50 is a 
top plan view of the vortex mixer with the vortex top removed. FIG. 51 is 
a side cross section view of the mixer showing the arrangement of heat 
transfer fins and the internal shape of the device with the top included. 
FIG. 52 is a side elevation view of the vortex mixer showing its mode of 
connection to a carburetor (or other fuel-air metering device) and to the 
engine intake manifold. 
See FIG. 50. The vortex comprises an integrally die cast housing 239. Flow 
through carburetor throttles (not shown) at 250 passes into chamber 251. 
Flow from chamber 251 into the main mixing section of the vortex is 
controlled by variable restriction throttle arrangement 254 which pivots 
on shaft 252, which is mounted directly so that its coefficient of 
discharge is linked to the coefficient of discharge of the carburetor 
throttle (not shown). Flow past the restriction forms wall attached 
streams which flow past fluid mechanically clean transitions into a vortex 
section which is basically a circular section around the outlet 242, which 
feeds the intake manifold of the engine. The outside peripheral walls of 
the vortex chamber, in interaction with deflector 256 are arranged so that 
a very significant fraction of the isentropic velocity past restriction 
254 is conserved so that the flow in the generally circular vortex section 
of the mixer has very high angular momentum with respect to the center of 
the outlet 242. In consequence, a strong irrotational flow towards the 
central sink 242 is formed. This flow is dominated by the physical effects 
of conservation of angular momentum as will be described below. For this 
flow system the tangential velocity of the flow is inversely proportional 
to the radial distance from the vortex center. Deflector vanes 240 are 
arranged at the outlet to recover a significant fraction of this 
tangential velocity back to pressure and to smooth the transition of the 
flow into the outlet, so as to very substantially reduce the total flow 
resistance of the vortex mixing device. Referring now to FIGS. 51 and 52, 
it can be seen that the outside peripheral wall of the vortex is heated by 
exhaust passage 268 extending around the outer walls of the vortex which 
includes heat transfer fins 270 and a closure member 272. The exhaust flow 
through the vortex can readily be arranged using a crossover arrangement 
analogous to the arrangements currently used for intake manifold heating. 
Consideration of FIG. 52 should make clear the positioning of the vortex 
with respect to the carburetor 282 (or other fuel-air metering device) and 
with respect to the engine intake manifold. Those skilled in the art of 
internal combustion engines will recognize that the vortex mixer of FIGS. 
50, 51, and 52 can be constructed as a simple and inexpensive die casting. 
Experimentally determined engine operating characteristics with the vortex 
mixer of FIGS. 50, 51, and 52 have been extremely excellent and the flow 
inside the mixer is an excellent approximation of the analytically 
predicted flow and mixing relationships. Because the fluid mechanics 
inside the vortex mixer illustrates structured flow turbulent mixing 
analogous to the structured flow turbulent mixing inside the cylinders 
disclosed in the present invention, it is worthwhile to discuss the fluid 
mechanics and mixing in the vortex mixer. A distinction between the vortex 
flow in the cylinder according to the present invention and the vortex 
flow in the vortex mixer should be emphasized. In the cylinder, an 
irrotational flow vortex has tangential velocity but no mean flow towards 
the center of the vortex (except for a recirculation due to boundary layer 
effects which will be small). In the vortex mixer, on the other hand, 
there is, in addition to the tangential flow, flow from the outside of the 
vortex to the center of the vortex. In consequence, the flow stream lines 
of the two vortex systems are appreciably different. However, the 
analogies in the reasoning involved in the flow inside the cylinder and 
the flow inside the vortex mixer should be clear. 
In the vortex mixer, conservation of angular momentum, MV.sub.t r, dictates 
the increase in the tangential velocity of the fluid as it flows towards 
the center. It is easy to verify that the velocity in the tangential 
direction as a function of radius r, V.sub.tr, will be expressible 
according to the relation 
##EQU1## 
where V.sub.to is the tangential velocity at the outside of the vortex, 
r.sub.o is the radius at the outside of the vortex, and r is the radius 
where the velocity tangential is taken. FIG. 53 illustrates the flow 
velocities which are produced in an irrotational flow vortex according to 
the above equation. 
Because the flow is flowing from the outside of the vortex to a sink at the 
center of the vortex, the mass flow rate in the radial direction through 
any cylindrical cut of the vortex section will be the same, so that the 
radial velocity will be inversely proportional to the radius 
##EQU2## 
where V.sub.rr is the radial velocity at radius r, and V.sub.rro is the 
radial velocity at the outside radius of the vortex. Clearly the above two 
equations are of the same form. It follows that for a set tangential 
velocity input (set by a given intake manifold vacuum) and a set volume 
throughput through the vortex (set for a specific rpm) the ratio of the 
velocity tangential to the velocity radial will be constant for all the 
radii of the vortex. 
It should be emphasized that the flow relations which have just been 
written down are not valid at radii less than the radius of the outlet. 
However, it should be clear that even though the irrotational vortex is 
not a perfect approximation of the flow, the physical relations of 
conservation of angular momentum make it conceptually rather close to the 
mean flow streamlines which do, in fact, occur and have the perturbations 
of turbulence superimposed upon them, if the boundary layer flows are 
properly controlled as they are in the vortex of FIGS. 50, 51, and 52. 
This boundary layer control, discussed in detail in the copending vortex 
patent application, will not be discussed here. It should also be clear 
that drag interactions between successive radial elements will tend to 
reduce the velocity increase of the flow as it flows towards the center, 
because the angular momentum as the flow flows towards the center will be 
reduced by these drag interactions. Nonetheless, the irrotational flow 
vortex form, as a flow mode, is extremely stable, and flow in the vortex 
mixer approximates this ideal fairly closely. 
As has been discussed previously in this application, the interaction 
between mean flow streamlines and turbulence must be understood if one is 
to understand mixing. A consideration of FIGS. 54, 55, and 56 should 
clarify some of the points important with respect to understanding of the 
interaction between flow structure and molecular and turbulent diffusive 
mixing in the vortex mixer. It should be emphasized that the graphical 
illustration of FIG. 54, FIG. 55, and FIG. 56 are exemplary only. However, 
the examples are important ones. FIG. 54 shows a streamline 354 of a 
clockwise rotating vortex from an outside radius 350 to a sink 352 where 
the streamline obeys the flow equations previously discussed. This flow 
streamline would occur, for example, in an irrotational flow vortex where 
the streamline was well away from entrance condition perturbations and 
where turbulence in the vortex was zero, if one were at point 356 to 
introduce, for example, ink into a water vortex and watch the ink line as 
it flows towards the sink. The streamline, in other words, shows what the 
flow path would be in the absence of any random mixing, either by 
turbulent diffusion or by molecular diffusion. If there were any 
diffusion, the width of the line would increase as it flowed towards the 
sink, as should be clear to those who understand mixing. In summary, FIG. 
54 would show a flow streamline for an irrotational flow vortex if a line 
of mixant was introduced at only one point along the outside of the vortex 
and in the absence of either molecular or turbulent diffusion. 
FIG. 55 shows what would happen if the same flow situation as that of FIG. 
54 had an additional line of mixant introduced 180.degree. around from the 
initial point of introduction. The vortex would have an outside circle 357 
and a sink 358. At point 360 along the circle 357 a line of mixant would 
be introduced 359. The numbers 359 are shown as the flow swirls in towards 
the sink to identify that streamline. 180.degree. from point 360 along 
circle 357 mixant is introduced at 362 and produces flow streamline 361. 
Flow streamline 361 is identified at several points to make it clear the 
manner in which the spiral 359 and the spiral 361 nest. Again, FIG. 55 
illustrates what would happen in a mathematically perfect irrotational 
flow vortex with a sink, in the absence of either molecular diffusion or 
turbulent diffusion. 
FIG. 56 is analogous to FIG. 55, except now, rather than having two nested 
spiral streamlines, mixant would be introduced evenly around 10 points 
around the circumference of the vortex; and therefore, 10 different spiral 
lines would nest as shown. 
With respect to FIGS. 54, 55, and 56, it should be clear that the presence 
of small-scale turbulent perturbations and molecular diffusion would tend 
to thicken out the lines as they flow from the outside towards the sink of 
the vortex and therefore that the mixing pattern would be more and more 
homogeneous as the mixture flowed towards the sink of the vortex. For 
example, with respect to FIG. 56, it should be clear that only a 
relatively small spreading angle of the mixant lines (corresponding to a 
relatively small turbulence intensity) would so smear out the lines of 
mixant by the time the flow had spiraled from the outside of the vortex to 
the sink, that the mixture at the sink of the vortex would be very 
homogeneous. With respect to the nesting of spiral streamline patterns 
shown in FIGS. 55 and 56, it should be pointed out that the fuel 
evaporation mixing of the vortex mixer shown in FIGS. 50-52 will come from 
fuel evaporating around the entire peripheral wall 260 of the vortex and 
that this will correspond to introduction of mixant not around 10 points 
around the periphery but around an effectively infinite number of points 
around the periphery. This means that if the liquid is well distributed 
around the circumference of the vortex peripheral wall (it has been 
determined experimentally that this is adequately near true), the mean 
distance across which diffusion needs to occur in order to achieve 
essentially perfect homogeneity at the vortex sink is very short. A 
consideration of the turbulent or molecular diffusion differential 
equation should make clear that an n-fold decrease in the mean distance 
across which diffusion needs to occur, for a set interfacial area, will 
decrease the time required for equilibrium by a factor of n. However, in 
addition, introduction of mixant from many points around the periphery of 
the vortex is tantamount to very vastly increasing the interfacial area 
across which diffusion can occur, and, of course, this effect increases 
mixing rates, too. Again, it must be emphasized that the flow streamlines 
shown in FIGS. 54, 55, and 56 are only exemplary. However, the geometrical 
relations with respect to mixing illustrated by these figures are 
extremely important and do not become less important as the flow 
structures become more complex; for any given flow structure, the flow 
structure will serve to stretch out the concentration gradients of species 
to be mixed and therefore, the flow structure will dramatically affect the 
rate at which mixing proceeds. Mathematically, the flow structure, or 
nonrandom streamline pattern, can be thought of as a spatial transform of 
concentration fields as a function of time as was discussed before. There 
are flow transforms which are very conducive to mixing, and the vortex 
mixer's irrotational flow vortex is such a flow transform. It should be 
clear that flow patterns which are not exactly irrotational flows can also 
have flow patterns very much conducive to mixing. For example, the flow 
pattern in the vortex of the present invention will not be a perfect 
irrotational flow vortex. With respect to the spiral streamlines, it will 
differ from a conventional irrotational flow vortex in that the ratio of 
tangential to radial velocity will not quite be constant as a function of 
radius for the real flow. However, nonetheless, Reynolds number modelling 
of the flow pattern produced by the present invention using water as the 
model and ink as tracer shows that the flow pattern which actually occurs 
in the system is much like an irrotational flow vortex and that the flow 
pattern is extremely conducive to mixing. In fact, when a single point 
mixant introduction (using ink from a syringe) was used, the mixing was so 
rapid that the flow looked effectively homogeneous well before the flow 
reached the outlet of the vortex. The ratio of tangential to radial 
velocity was about constant over various radial distances, verifying that 
the flow was basically an irrotational vortex. Consideration of the flow 
nesting relations in FIG. 54, FIG. 55, and FIG. 56 should make it clear 
that the mixing must have been even better for the multiple mixant 
introduction case in the real vortex where fuel is distributed for 
evaporation around the circumference of the outer wall of the vortex. 
While it is recognized that viewing a Reynolds number model operating on 
water with ink as a tracer in a plexiglass one-to-one model is not quite a 
perfectly analogous modelling (because the water is not compressible as 
the air is), the analogy is still a close one, and the mixing observed in 
the system was very, very intense, so that even significant decrements in 
mixing rates due to compressibility effects (which are not likely) would 
not affect the conclusion that the vortex flow pattern actually produced 
in the vortex mixer system is extremely conducive for mixing, so that with 
the multiple fuel point evaporation characteristic of the evaporation 
process of the vortex mixer, the mixture at the outlet will be effectively 
homogeneous. 
The theoretically excellent performance of the vortex mixer shown in FIGS. 
50, 51, and 52 has been verified by a number of experiments. In a Ford six 
cylinder engine equipped with a vortex mixer, cylinder to cylinder 
variation was measured with a Horiba air-fuel ratio meter, and cylinder to 
cylinder variation was undetectable under all the conditions tested. The 
air fuel ratio meter should have been able to resolve an air-fuel ratio 
variation of a tenth of a ratio. The Air-fuel ratio meter, however, has 
relatively slow response and therefore the transient characteristics of 
the system were measured as follows. 
A Ford five liter eight cylinder engine was equipped with oxygen sensors at 
each exhaust port, in addition to an oxygen sensor collecting off one four 
cylinder bank of this engine. The oxygen sensors were the same oxygen 
sensors now used for three-way catalyst controlled engines. These sensors 
can be made into extremely fast acting and sensitive devices by removing 
the protecting diffusion barrier shroud around the ceramic sensor element, 
leaving the ceramic sensor surface directly in the flow. When this is 
done, the tau (1/e response) of the sensor when hot appears to be in the 
vicinity of 5 milliseconds. It is believed that when the oxygen sensor is 
exposed to a truly homogeneous exhaust mixture (on the microscale), it 
will switch from a low voltage to its full output voltage in something 
less than one tenth of an air-fuel ratio when the oxygen sensors are hot. 
The full-scale response of the oxygen sensor is therefore expected to 
represent less than a one percent variation in air-fuel ratio. With 
present production techniques, oxygen sensors are not yet quite identical, 
so that absolutely perfect comparisons of oxygen concentration and 
air-fuel from cylinder to cylinder cannot at present be made. However, use 
of the oxygen sensors makes it possible to verify the tightness of 
cylinder-to-cylinder distribution, and since the oxygen sensor cannot 
respond to individual exhause stroke oxygen concentrations, the oxygen 
sensors also give an excellent read on cycle-to-cycle mixing statistics. 
For these reasons, the inventor and his associates Kenneth W. Kriesel and 
Charles L. Siewert adapted the Ford five-liter engine to have oxygen 
sensors directly downstream of each intake port (in a position where the 
oxygen sensors operated very hot increasing their sensitivity and reducing 
their full-scale switching range). In addition to the eight oxygen sensors 
corresponding to the eight cylinders, an additional oxygen sensor at the 
outlet of one of the exhaust manifolds was added. In this way, 
cylinder-to-cylinder and cycle-to-cycle variations which could in no other 
manner be measured were measurable. FIG. 57 shows actual engine test data 
obtained with these oxygen sensors under a condition corresponding to a 
very heavy accel on the EPA cycle and with the carburetor operated without 
any accelerator pump. In the chart trace, the upward direction is the 
direction of increasing time. Heavy, horizontal automatically written grid 
lines represent seconds and the lighter horizontal grid lines represent 
tenths of a second. The right most trace, trace 10, is the trace for 
voltage on the servo-electric linkage used to actuate the carburetor 
throttle during these tests. The one-second period when the throttle was 
opened in the specified way is shown on the trace. Proceeding from right 
to left are traces for cylinders 1, 5, 4, 2, 6, 3, 7, and 8 and the left 
most trace is the average trace for cylinders 5 to eight (the trace from 
the oxygen sensor at the collector of the left exhaust manifold). The 
temperature, and therefore the output of the cylinder 5 to 8 average trace 
are different from those of the oxygen sensors directly in the exhause 
ports of the various cylinders. The test was conducted with the air-fuel 
ratio from the carburetor prior to the servo-controlled opening of the 
carburetor throttle adjusted so that the oxygen sensors were on the rich 
side just at the edge of switching (in practice operated so close to the 
switching point that the switching noise on the oxygen sensors were 
clearly seen). During an accel with a carburetor unequipped with an 
accelerator pump, it is expected that the fuel will lag the air and 
therefore that the engine will suffer a lean excursion. With usual engine 
arrangements this lean excursion for the conditions of test for FIG. 57 
would constitute several air-fuel ratios and the lean excursion would last 
for several seconds (and involve quite noticeable decrements of engine 
performance). The test illustrated in FIG. 57 was conducted with the Ford 
engine equipped with a mixing vortex such as that shown in FIGS. 50, 51, 
and 52. After the carburetor throttle began to open on this test, there 
was a lean excursion, but it was of the order of a tenth of an air-fuel 
ratio or less (significantly less than one percent). Also looking at the 
traces of the various sensors, it is clear that the variation is little 
different from one cylinder to the next (particularly in view of the fact 
that the oxygen sensors themselves are not identical). This data, and much 
data like it, verifies that the cylinder-to-cylinder and cycle-to-cycle 
mixing statistics of the vortex are much tighter than a standard deviation 
of one tenth of a percent. The mixing statistics achieved by the vortex 
mixer are significantly better than those of any other mixing system, of 
any type or at any price, of which the inventor is aware. Experimentally, 
it was found that at present the major source of statistical variation in 
air-fuel delivery to the cylinders was due to statistical variations in 
the metering characteristics of the carburetor itself. Improvements in 
fuel-air metering systems are proceeding rapidly in many laboratories, and 
it is expected that very tight fuel-air metering, particularly from a 
single point fuel introduction arrangement can be commercially achieved. 
In all events, the vortex mixer is a very straightforward way to achieve 
the tight cycle-to-cycle and cylinder-to-cylinder mixing statistics which 
when combined with excellent in-cylinder mixing and turbulent flame speed 
control, permit ultra-low nitric oxide emissions with optimal fuel 
economy. 
SINGLE CYLINDER TEST DATA VERIFIES NO.sub.x AND FUEL ECONOMY ADVANTAGES OF 
THE INVENTION 
Automotive engineering requires, for clear commercial reasons, very high 
standards of proof, particularly with respect to technology concerned with 
so sensitive a subject as emissions. The present application has been 
detailed and lengthy because it contains the information required to teach 
skilled automotive engineers, who are not commonly acquainted with 
fluidics and have difficulty visualizing and thinking about structured 
flows and the mixing interactions related to structured flows, how to 
understand, make, and use fluidic ports in combination with homogeneous 
lean mixtures to produce very efficient and ultralow NO.sub.x output 
engines with improved peak power and excellent driveability. Up to this 
point, the invention has been explained largely in light of quite complex 
interrelated theoretical arguments, with relatively little test data 
specifically considered. Any automotive engineer with experience knows 
that theory is not enough. Also it is very well established that theory, 
even very carefully contrived theory, very often founders in practice due 
to effects inadequately understood by the theoretician. Because of many 
negative experiences in the application of theory to practical engines, 
the automotive engineering profession tends to be quite dubious when 
confronted with theoretical arguments, even when these arguments are 
short. In the present case, the arguments and theoretical bases of the 
present invention are quite complex and interrelated. It is an elementary 
proposition of practical logic that an argument is the more suspect the 
longer and the more complicated it is. Also, an argument is rightly held 
to be suspect when its conclusions are at substantial variance with what 
is commonly believed in the discipline to which the argument relates. For 
these well founded reasons, the arguments of the present inventor, 
considered as logical propositions, were not obvious or considered as 
adequate bases for action by men of experience, responsibility, and power 
in automotive engineering. For these men, skilled in the art of internal 
combustion engines, it can be said that no logical proposition, no matter 
how carefully crafted and no matter how carefully related to other engine 
data, is obvious. The tradition of automotive engineering is, and has been 
for some decades, a tradition of rigid empiricism. For these reasons, the 
inventor's effort to secure support from major auto companies on the 
subject matter of the present invention was unsuccessful, even after a 
rather substantial body of data relating to the fluidic port variable 
restriction engine was available. At the same time when the major auto 
companies were spending several billions of dollars a year on emission 
control research and development, the inventor's work was held to be 
unworthy of support, in the face of extremely complete (and in the event 
correct) technical arguments and much fluid mechanical data. This was true 
even though the inventor was given no technically coherent argument on why 
his invention would not work as he expected, and even though several 
engineers agreed that the invention, if it operated as expected, would be 
by a large margin superior to the technology on which these major 
companies were working. The inventor's support level from the academic 
community was significantly greater, but still highly tentative. The plain 
fact was that low NO.sub.x, high efficiency results which the inventor 
held to be possible and which motivated the inventor's work were held to 
be impossible by an overwhelming consensus in the automotive engineering 
profession. The only answer to such a consensus must be data, and the data 
must be accurate and conclusive beyond question. The inventor therefore 
had to accumulate a great deal of data at the Internal Combustion Engine 
Research Laboratory of the University of Wisconsin under the quite close 
supervision of Professors P. S. Myers and O. A. Uyehara. The support of 
Professors Myers and Uyehara was significant, since without resources 
under their control, the research could not have been conducted. Myers and 
Uyehara were also a very significant disciplining force on the inventor 
and worked very hard to assure themselves that the inventor drew no 
overoptimistic conclusions from his work. At this time, Myers and Uyehara 
had great practical power over the inventor, since they would have quickly 
disowned his work and so discredited him, if he had presumed to say 
something was true which they had not themselves seen proved beyond any 
reasonable doubt. 
Because of the controversial nature of the research, the single cylinder 
engine test facility set up to test the variable restriction port engine 
development was built to uncommon standards of accuracy. A very great deal 
of novel technical work was done in order to secure the very high 
standards of accuracy needed. A special hydraulic torque sensing 
arrangement where torque was read on a mercury manometer, where damping 
was strictly linear, and where all sources of static friction were 
essentially eliminated was built. It is believed that this torque meter 
was at all times accurate to within plus or minus one tenth of one 
percent, which is very substantially superior accuracy to the accuracy 
available with more conventional torque measuring means. Engine speed was 
measured with the conventional sixty tooth gear and one second electrical 
timer apparatus to give a digital readout of engine r.p.m. Air flow 
metering was accomplished using critical flow orifices and specially 
calibrated air gauges, with special compensation of air temperature to 
compensate for effects of temperature variation on air supplied to the 
test setup. This air metering arrangement was tested and found to be 
within plus or minus 0.2 percent, which is again substantially more 
accurate than the conventional air metering setup held acceptable for 
conventional engine research. Fuel flow metering for gaseous fuels was 
also accomplished using critical flow orifices (watch jewels) with 
specially calibrated gauges and with a heat exchange arrangement which 
assured that the temperature of the fuel gas upstream of the critical flow 
orifice did not vary by more than approximately 1/2.degree. C. For the 
propane metering arrangement, reproducibility and accuracy in the vicinity 
of +/-0.1 percent was attained and measured with the most accurate 
volumetric measuring apparatus available at the University of Wisconsin. 
The fuel metering apparatus was very much more accurate than that 
conventionally used for automotive engine tests. Intake manifold vacuum 
and other relevant near atmospheric pressures were measured with mercury 
manometers. The effect of pressure fluctuations on the manometers was 
filtered by laminar flow elements in each pressure pickup line, so that 
manometers measured true linear averages for the fluctuating pressures. 
Pressures inside the cylinder were monitored with a Kistler pressure 
probe, but this probe was not precisely calibrated since the research did 
not require such calibration. Emission tests were taken with the emission 
measuring cart available at the Internal Combustion Engine Research 
Laboratory. NO.sub.x was measured with a chemiluminescent analyzer by 
Thermo-Electron. Hydrocarbon emissions were measured with a Beckman flame 
ionization flame detector hydrocarbon meter. Oxygen was measured with a 
Scott Research oxygen meter as a check on fuel-air ratios. Carbon monoxide 
was measured with an infra-red absorption meter; however, in the lean 
range of the present research the carbon monoxide concentrations proved to 
be too small to be reliably measured with this instrument (and too small 
to be of practical concern). The research was conducted on an electric 
dynamometer with motoring capability, and was done on a single cylinder 
engine set up on a CFR crankcase. The engine used for the tests was a 
one-to-one one-cylinder model of a 1951 Oldsmobile engine having a bore of 
3.5 inches and a stroke of 3.75 inches and a compression ratio of 8.3 to 
one. This engine, although obsolete, was the only single cylinder test 
engine available to the inventor which had the tangential ports with 
respect to the cylinder center which are characteristic of multiple 
cylinder engines and which are required for the proper function of the 
fluidic port. 
The single cylinder engine used was not ideal, although it was sufficient 
to establish a great deal of useful data. The main problem with the engine 
was that it was characterized under all operating conditions with unburned 
hydrocarbon emissions very much in excess of those of more modern engines. 
Another problem was that the intake ports of the engine tested were not 
quite optimal. However, the inventor had little choice but to use the 
heads available and, of course, could do a less adequate job of port 
modification with welding and filing than he would have been able to do 
had he been permitted to vary port cores on an original casting. 
Nonetheless, the fluidic ports did function well, although the fluidic 
efficiency of the ports on the test engine was always significantly below 
the fluidic efficiency of the intake ports which the inventor had modified 
from a 350 cubic inch displacement Chevrolet engine head. 
Many of the data points taken were taken to find the optimal performance 
with respect to fuel economy for the engine holding certain variables 
constant. For the technique, the inventor is indebted to the optimization 
process described by Professor Paul H. Schweitzer, who was also a 
significant teacher for the inventor. In this optimization process, the 
fuel flow is set constant, the r.p.m. of the engine is set constant, and 
the other variables (except perhaps some specifically held constant) are 
varied so as to maximize the torque (hence the horsepower) of the engine. 
Mathematically, what is done in the Schweitzer optimization process is to 
take the partial derivative of horsepower holding fuel flow and r.p.m. 
constant and varying the other variables. In this way, that combination of 
fuel-air ratio, spark advance, EGR, and in the case of the present 
invention flap setting which gets the maximum power from a given fuel flow 
can always be found. It should be pointed out that the slope of the power 
curve very near the optimal is very flat, and so one must measure torque 
very accurately in order to accurately determine the actual optimal point. 
In addition, since a number of variables must be changed, the discovery of 
an optimal point is a relatively time consuming experimental process, so 
that each optimal point on a graph represents the end point of a 
relatively long experimental sequence. Nonetheless, a great many of the 
following data points are the result of the optimization process because 
it was felt that this offered the toughest and ultimately the fairest test 
of the data. Using the optimization process, it has been shown that the 
conventional trade-off between nitric oxide emissions and fuel economy 
disappears if the fuel-air mixing process is sufficiently good. 
The first single cylinder experiments were done with the engine adjusted to 
a 10:1 compression ratio and operated on methane fuel introduced into a 
mixing tank. FIGS. 58 and 59 show representative results of this engine. 
The results made clear that the variable restriction fluidic port 
significantly improved efficiency and shifted the optimum fuel economy 
mixtures significantly leaner, as the inventor had expected. During that 
time when the date of FIG. 58 and 59 was taken, the engine was not set up 
for emission measurements. 
One important difference between the data of FIGS. 59 and 59 and the lower 
compression ratio data to follow was that the fraction of fuel-air cycle 
efficiency shown with the high compression ratio engine was significantly 
higher than that shown with the lower compression ratio engine. The 
inventor believes that the explanation is as follows. With the high 
compression ratio engine, there was significant squish around top dead 
center piston position, and this squish distrupted the flow structure 
inside the combustion chamber, reducing heat losses during the power 
stroke. With the lower compression ratio piston, tdc clearances were great 
enough that squish was insignificant, and this is believed to be the 
explanation for the higher heat losses of the lower compression ratio 
engine. If this explanation is correct, squish is highly desirable for a 
fluidic port equipped engine. However, the inventor has not directly 
verified that squish is useful for reducing heat losses, although the 
effect seems very likely. 
The experimental sequence which produced FIGS. 58 and 59 made clear that 
the test results that the inventor had gotten with his full size engine 
were not a fluke, and that, indeed, the best economy air-fuel ratio could 
be extremely lean, and the efficiency could be improved with improved 
in-cylinder mixing. However, Professors Uyehara and Myers felt that 
methane was a somewhat unrepresentative fuel (although easy to meter, 
since it was an ideal gas) and suggested that the test setup be modified 
to run with propane, a more typical fuel and a fuel where automobile 
companies have accumulated large masses of test data. The data points 
plotted in the following figures were accumulated with the test engine 
operated on propane. 
DRASTIC REDUCTIONS IN NO.sub.x AND IMPROVEMENTS IN FUEL ECONOMY WITH THE 
VARIABLE RESTRICTION ENGINE 
The effects of equivalence ratio and mixing on NO.sub.x formation in 
engines have been discussed in some detail with respect to FIGS. 3, 4, 5, 
and 6. In addition, very advanced computations, too complicated to be 
considered in a patent application, were conducted for the inventor by Dr. 
J. Carl Pirkle (then of the Johns Hopkins University and now of Exxon 
Research and Engineering Corporation) which integrated the kinetic 
equations characteristic of engine operations in enough detail to make 
quite sure that very low NO.sub.x results were possible with extremely 
dilute combustion. These calculations by Dr. Pirkle were a strong 
encouragement to the inventor to persevere in his work in the face of many 
discouragements. It should be said that a significant body of the engine 
literature evidence stood against the inventor in his belief that 
ultralean engine operation with very good mixing would permit the ultralow 
NO.sub.x results he was seeking. It was, for example, well established in 
the literature that at the normal lean limits of engine operation there 
were no great effects of additional mixing with respect to NO.sub.x 
output. In addition, the inventor had been told by authoritative 
researchers actually working in the field that the degree of in-cylinder 
turbulence and swirl required in the present invention would result in 
heat losses such that any thermodynamic advantages of the system would be 
outweighed by the additional heat losses. In addition, the inventor was 
setting out to burn mixtures so dilute they were, according to the 
conventional wisdom, substantially incombustible under engine operating 
conditions. 
FIG. 60 shows very important experimental results which verify the basic 
theory and practice of the invention, and which also show why the 
advantages of superhomogeneous combustion would not be apparent to one who 
was extrapolating engine data in the conventional way. The data in FIG. 60 
was taken at a set engine speed of 1200 r.p.m. with the fuel flow set at a 
constant rate of 1.30 pounds of propane per hour. For each data point 
spark advance was set at the MBT setting (as will be shown, this 
significantly understates the advantage of the present invention with 
respect to more conventional spark settings) and because of the very 
important nature of the results shown in FIG. 60, many of the data points 
were repeated. Because of the requirement to establish MBT spark timing 
within at least one degree, the data in FIG. 60 represents a significant 
amount of experimental time. It is believed, and it will be made clear 
later, that the bulk of the scatter in the data of FIG. 60, even with 
careful determination of spark timing, is due to statistical variations in 
the MBT spark timing determination from point to point. The experimental 
sequence for taking the data was as follows: The engine was fully warmed 
up (about the power of steady state running) at roughly the stoichiometric 
ratio; during this time the NO.sub.x measuring apparatus was also warmed 
up and calibrated when fully warm (the calibration of the NO.sub.x meter 
was rechecked after each experimental day, and drift was found to be 
neglible). For the data points marked wide-open flap port injected, the 
variable restriction flap of the engine setup was fully open, so that the 
engine produced a flow pattern not too different from that of a 
conventional stock engine (although with a irrotational flow component 
which was necessitated by the fluid mechanics of the particular port 
used). The propane fuel was injected approximately six inches ahead of the 
intake port itself, so that there was significant heterogeneity of the 
charge inducted into the engine. It is believed that the degree of 
heterogeneity characteristic of this wide-open flap port injected run is 
fairly characteristic of the heterogeneity actually delivered in 
conventional engines under normal operating conditions. For the data 
points marked with X's the variable restriction flap was closed until only 
a 0.300 inch gap (measured with a very precise variable restriction 
Vernier arrangement) was open between the floor of the port and the flap 
restriction which was a flap restriction similar to that shown in FIG. 
34B. For the 0.300 inch flap opening port injected points the port 
injection arrangement was identical to that of the wide-open flap port 
injected run. For the data points marked with triangles the flap opening 
was again 0.300 inches, but the mixture was premixed with a vortex (which 
was not heated since propane was already a gaseous mixture) prior to 
delivery to the engine. It is believed that the mixture from the cold 
vortex mixer was substantially homogeneous prior to introduction to the 
engine intake port, although the vortex mixer used in these experiments 
was substantially more primitive than that which has previously been 
discussed. 
FIG. 60 graphs the variation of NO.sub.x output versus equivalence ratio 
with the NO.sub.x output expressed as grams NO.sub.2 per indicated 
horsepower hour, on a logarithmic scale, since the NO.sub.x output varied 
by more than a factor of a thousand from point to point. The abscissa of 
the graph is equivalence ratio. Relationships between equivalence ratio 
and the more typically used air-fuel ratio are clearly shown on the 
abscissa of FIG. 3 (which was taken from a publication of Paul N. 
Blumberg). 
In FIG. 60 it is shown that at the relatively rich ratios characteristic of 
conventional engine operation, there is not much advantage to the extreme 
homogeneity supplied by the present invention. Between an equivalence 
ratio of 1.0 and 0.8 stoichiometric, the difference in NO.sub.x output 
among the wide-open flap port injected case, the 0.300 inch flap port 
injected case, and the 0.300 inch flap cold vortex mixer case is neither 
particularly large nor very convincing. However, in the range leaner than 
0.8 equivalence ratio, the situation changes drastically. For the no port 
restriction case, the best fuel economy happens at an equivalence ratio of 
0.735, so that enleanment beyond this point results in quite significant 
fuel economy penalties. From the equivalence ratio at the very leanest 
limit where engine operation was possible (0.59 equivalence ratio) to the 
richest mixture tested at stoichiometric, there is only a little more than 
a factor of ten change in the NO.sub.x output in the conventional case. 
Furthermore, most of this NO.sub.x range involves a fuel consumption 
penalty: the optimal indicated specific fuel consumption level for the 
conventionally set up engine involves NO.sub.x outputs of almost half the 
maximum NO.sub.x outputs for this engine under these conditions. It is 
results like those shown here for the conventionally set up case which 
have convinced the industry that NO.sub.x control via charge dilution is a 
relatively unattractive procedure incapable of producing the very low 
NO.sub.x output levels which will be required by the federal government. 
It should also be said that the relatively low reductions in NO.sub.x 
level with enleanment are anomalous with respect to chemical kinetic 
calculations in engines, and that the industry has largely discounted 
these chemical kinetic calculations in the face of this data. 
In the range leaner than 0.8 equivalence ratio, the NO.sub.x performance of 
the variable restriction engine is drastically better. Quite clearly, for 
a set equivalence ratio leaner than 0.8 equivalence ratio, the variable 
restriction engine with its superior in-cylinder mixing had significantly 
lower NO.sub.x levels. For example, for the 0.300 inch flap restriction 
port injected case, the optimum fuel consumption (minimum indicated 
specific fuel consumption point) occurred at an equivalence ratio of 0.59 
equivalence ratio at a NO.sub.x level of 0.27 grams per indicated 
horsepower hour. This represents a 48 fold reduction in NO.sub.x level 
from the maximum for the variable restriction engine, in contrast to only 
a 2.4 fold reduction in NO.sub.x output from the maximum for the 
conventional engine setup when comparing the fuel consumption optimal 
NO.sub.x output to the maximum NO.sub.x output. In addition, the indicated 
specific fuel consumption with the variable restriction engine was 
significantly and reproducibly less than the indicated specific fuel 
consumption of the optimal setting for the conventionally set up engine, 
so that the drastic reduction in NO.sub.x was attained with a simultaneous 
(although relatively small) improvement in the fuel economy of the engine. 
The advantages of homogeneity and structured flow turbulence are even more 
dramatically shown for the case where the variable restriction engine is 
supplied with a mixture homogenized by the vortex. In this case the 
optimum indicated specific fuel consumption is even lower, and the 
reduction in NO.sub.x from the maximum is a very large factor of 420. The 
relationship between equivalence ratio and NO.sub.x characteristic of the 
variable restriction setup engines is extremely close to that predicted by 
the chemical kinetic calculations of Pirkle and others and stands in 
significant contrast to the kinetically anomalous NO.sub.x versus 
equivalence ratio characteristic of the conventional engine. The 
anomalously small reductions in NO.sub.x with mixture dilution for the 
conventional engine setup are due to inadequate mixing. Several points 
with respect to the data shown in FIG. 60 should be emphasized. The data 
shows the effect of mixture dilution via enleanment but does not show data 
employing exhaust gas recirculation because of the great practical 
difficulty in metering exhaust gas flows to an engine to the level of 
accuracy characteristic of the inventor's test setup. Those skilled in the 
internal combustion engine arts, and those skilled in chemical kinetics 
will recognize that exhaust gas recirculation is generally considered to 
be a more effective diluent than excess air, due to the fact that flame 
stability is generally better with high EGR than with very lean mixtures 
having the same NO.sub.x output. There is every reason to expect that the 
variable restriction engine would show the same qualitative and 
quantitative advantages with respect to NO.sub.x reduction and improved 
fuel economy if EGR was employed as a significant part of the diluent; 
however, it is expected that the absolute value of the NO.sub.x levels 
with exhaust gas recirculation would be even lower than those shown here. 
It is also believed that there is still room for considerable improvement 
in the in-cylinder mixing rate of the engine. The fluidic performance of 
the intake port of this test engine was significantly inferior to the 
fluidic performance of prototype port designs which the inventor has 
modified from a 350 c.i.d. Chevrolet head, and it is expected that with 
improved port fluidic efficiency improved mixing would be attainable. A 
reader who has carefully considered the fluidic information in this case 
will recognize that a significant amount of careful experimental work will 
be necessary to produce an absolutely optimal flow structuring in any 
engine. Nonetheless, if the fluidic ports do produce the tangential flow 
disclosed in the present case, a structured flow in the engine having 
excellent flow results will occur. It is expected that it will be 
relatively easy to produce fluidic results superior to those of the test 
setup generating the data described here for any production engine. 
A very important point which can be seen by looking at FIG. 60 is the 
extremely rapid rate at which NO.sub.x output falls off in the very lean 
area. In this very lean range, a few percent variation in equivalence 
ratio can account for an order of magnitude change in the NO.sub.x output. 
This result is extremely reasonable kinetically. The result also has very 
practical implications: In the very lean range there is a very large 
payoff to widening flame stability limits, and even a relatively small 
change in flame stability limits can result in a substantial reduction in 
the NO.sub.x output of the engine at best fuel economy trim. Another point 
which is very clearly emphasized by FIG. 60 is that there is in the lean 
regime a very substantial payoff to mixture homogeneity. 
FIG. 60 shows very dramatic reductions in nitric oxide output, in addition 
to improved fuel economy with the variable restriction homogeneous charge 
engine of the present invention. With the fluidic port engine, NO.sub.x 
outputs a factor of a hundred less than those attainable with the 
conventional engine are available. The fluidic port structured flow mixing 
arrangements of the engine of the present invention represent a 
breakthrough in NO.sub.x control and drastically change the conventionally 
understood tradeoff between fuel economy and NO.sub.x to a point where the 
optimal setting from the point of view of fuel economy also produces 
ultraflow NO.sub.x emissions. It should be emphasized that the load of the 
engine for FIG. 60 varied between 61 and 69 p.s.i. indicated mean 
effective pressure, that the indicated mean effective pressure was at its 
maximum at the optimal indicated specific fuel consumption points, and 
that the indicated mean effective pressure of 69 p.s.i. is a relatively 
heavy load for conventional driving. A consideration of the data of FIG. 
60, in combination with the data to follow, should make the detailed and 
complex arguments of this case with respect to FIGS. 3, 4, 5, and 6 more 
meaningful, and should serve to verify a close relationship between the 
theory put forth in that discussion and the actual results in the engine. 
FLUIDIC ENGINE DATA CORRELATED AGAINST THE EPA-CVS EMISSION CYCLE 
Emission control in practice must be achieved on full size vehicles capable 
of flexible operation under all the many conditions where automobile 
engines are expected to behave well. However, setup of an engine in a 
vehicle with a particular control strategy is an extremely expensive 
process, and it has therefore become well established to predict the 
emission control characteristics of an engine control strategy from 
correlations before the actual effort of engine buildup is undertaken. In 
this way, many more strategies can be evaluated than would otherwise be 
possible, and the tradeoffs between one engine characteristic and another 
become more clear. One of the better correlations schemes between steady 
state engine performance and cycle performance in a vehicle is the 
correlation scheme of Paul N. Blumberg which is used by the Ford Motor 
Company (Powertrain Simulation: A Tool for the Design and Evaluation of 
Engine Control Strategies in Vehicles, SAE Paper 760158) In Blumberg's 
correlation procedure, the cycle is approximated by a matrix of speed-load 
points, with each point weighted so that the sum of the emission outputs 
from these points should predict the emission output of a well set up 
vehicle, using the control strategy evaluated with the correlation matrix. 
Since a full size engine buildup with the improved fluidic ports has not 
been completed, the correlation scheme offers a useful check on the 
emission control advantages which could be expected from the present 
invention engine. Blumberg's correlation scheme was originally set up 
plotting brake mean effective pressures versus engine speeds; it is well 
established that single cylinder engines have anomalously high friction 
when compared with multi-cylinder engines. To correct for this friction 
effect, the indicated mean effective pressure from the test engine was 
determined using the motoring method, and a friction mean effective 
pressure of 20 p.s.i. under all conditions was assumed. This friction mean 
effective pressure estimate was used because it was held to be an 
extremely conservative (pessimistic) estimate of the advantages attainable 
with the fluidic port engine. Those skilled in the internal combustion 
engine arts will recognize that the fuel consumption predictions from any 
such correlation scheme will be extremely sensitive to the value of 
friction mean effective pressure assumed. With a lower friction mean 
effective pressure assumption, predicted fuel economy would be 
significantly improved and NO.sub.x emissions would also be reduced. 
The Blumberg correlation matrix which the inventor was able to obtain 
applied to a 1972 Mark IV vehicle with a curb weight of 5500 pounds, which 
is drastically heavier than the weight of most new vehicles, and with a 
351 cubic inch displacement engine. Clearly, the predicted emission levels 
would be less for a lighter vehicle. 
FIG. 61 shows the indicated mean effective pressures and rpm's for the 
Blumberg 8 point matrix approximation of the Environmental Protection 
Agency CVS hot cycle which will be applied in the data immediately 
following. Blumberg was able to show that the correlation procedure when 
applied to Ford data, produced an excellent prediction of actual vehicle 
performance. Note that points are labeled by number. 
FIG. 62 plots NO.sub.x emissions in total gram contribution for the cycle, 
for each of the eight points on the Blumberg matrix. Comparing the 
Blumberg data points with the points for the maximum fuel economy setting 
for the variable restriction engine, using a mixing arrangement much 
inferior to that of the vortex (It is expected that had the vortex been 
used on this data, NO.sub.x levels from the fluidic port engine would have 
been reduced by something of the order of a factor of two). Note that the 
great bulk of the nitric oxide emitted over the cycle is represented by 
points 4, 5, 6, 7, and 8. For each of these points, the NO.sub.x reduction 
of the fluidic port engine when compared with the base line Ford 351 cubic 
inch engine is dramatic. The advantage is even greater than it appears, 
since the data for the fluidic port engine was done with MBT spark timing, 
while the spark timing on the '74 Ford engine involved a significant 
amount of spark retard. It should also be emphasized that the NO.sub.x 
reduction of the fluidic port engine was attained without benefit of 
exhaust gas recirculation. FIG. 63 plots in graphical form the NO.sub.x 
emissions in grams per mile NO.sub.2 which would be predicted for the 
fluidic port engine using the Blumberg eight point correlation running the 
engine on propane for fuel metering accuracy. In FIG. 63 is also plotted 
the very drastic relation between NO.sub.x output and spark timing, and 
also the strong relation between NO.sub.x output and equivalence ratio. 
For the absolute optimal equivalence ratio and MBT spark timing, the 
NO.sub.x output from the vehicle was predicted to be 0.313 grams NO.sub.x 
per mile. For the same air-fuel ratio, with a spark retard of 5 degrees, 
the NO.sub.x output was predicted to be less than 0.1 grams per mile 
NO.sub.x. The 5 degrees retard involved an approximately 2 percent penalty 
on fuel consumption and in exchange produced more than a three-fold 
reduction in NO.sub.x output. The graph also shows the very important 
effect of changing equivalence ratio. For the power optimal setting, if 
spark advance is held constant and the mixture is then set 10 percent 
richer than its optimal setting, the NO.sub.x output of the engine 
increases by essentially a factor of ten. A similar effect is present for 
the retarded spark cases. Clearly, the NO.sub.x output of the engine 
depends to a very great extent on the spark timing and the air-fuel ratio 
supplied to the engine. These points will be discussed in more detail in 
subsequent figures. 
FIGS. 64, 65, and 66 discuss the HC emissions from the engine, which are 
high. Although the test engine which was employed always had HC emissions 
more than a factor of two in excess of those of more recent engines, the 
HC emission output from the engine are still relatively high. The HC 
emissions, however, are in the catalyst controllable range and are, for 
example, in the same range as those of the Ford Proco engine which is 
currently under very heavy development. FIG. 64 for HC is analogous to 
FIG. 63 for NO.sub.x. From FIG. 64, it can be seen that the HC emissions 
at the power optimal are significantly higher than those for a 10 percent 
richer mixture, and one can see the common effect of spark retard for 
reducing HC emissions. FIG. 65 shows the HC emissions of the variable 
restriction engine on the Blumberg eight point matrix compared with the 
data for the 1974 Ford 351 baseline engine. The HC emissions with the 
variable restriction engine are significantly higher. It is not known to 
what extent the HC penalty is due to enleanment and to what extent the HC 
penalty is due to the very dilute combustion of the present invention. 
However, the HC increase with the very lean combustion was less than a 
factor of two over the HC output of this same test engine when adjusted to 
minimize HC. This is a relatively small percentage HC penalty for 
enleanment, when compared to the HC penalties more commonly encountered. 
However, the extent of the HC problem with the present invention has not 
yet been definitively established, and, although experimental work to 
control engine-out HC with this engine is at the conceptual stage, actual 
experimental work has not been undertaken. 
FIG. 66 answers the question, "How does HC emission vary with the setting 
of the variable port restriction, when other variables are held constant?" 
In the run plotted in FIG. 66 air flow and fuel flow were both set 
approximately constant, spark advance was held constant, and the flap 
setting was changed. There are two points which are plotted which could 
not later be reproduced and therefore have question marks; however, these 
points are included for completeness. An opening of 1.55 inches was the 
full open flap position, and the minimum flap opening position plotted was 
the minimum opening which was possible without changing air-fuel ratio by 
restricting air flow. It can be seen that the total range of indicated 
specific HC emissions for this run is between approximately 7.8 and 6.25 
grams per indicated horsepower hour hydrocarbon, which is a relatively 
small range. One could, perhaps, gather from the data of FIG. 66 that the 
variable restriction has a small but positive effect on HC emissions. 
However, the most conservative conclusion, for the equivalence ratio of 
the run of FIG. 66, is that the variable restriction has no significant 
effect on HC emissions. In all events, it is quite clear that the variable 
restriction does not make HC emissions worse. 
A very important issue with respect to dilute combustion is combustion 
smoothness. The commercial acceptability of an engine is not simply a 
matter of emissions and fuel consumption: The driver must feel that the 
engine is smooth and responsive and he will respond very badly to a rough 
running engine. FIG. 67 plots the variation in peak pressure measured with 
the Kistler pressure probe for the eight point Blumberg matrix. The data 
was taken by looking at peak pressures recorded in oscilloscope 
photographs taken under the operating conditions of the Blumberg optimals 
which have been discussed previously. The peak pressure variation of the 
engine was not zero, but it was well within the range typical of 
commercial engine practice. The peak pressure variations recorded 
correspond to less than a one percent variation in torque from cycle to 
cycle. It is expected that with a fluidically superior port arrangement 
these statistical variations in peak pressure could be further reduced. 
FIG. 68 records experimental data set up to answer the question, "What is 
the effect of the variable restriction flap opening on flame speed?" Those 
skilled in the internal combustion engine arts recognize that spark 
advance is a convenient variable to estimate flame speeds. The lower the 
best torque spark advance, the faster the flame speed. As can be seen from 
FIG. 68, as the flap restriction begins to close, there is a significant 
reduction in MBT spark advance initially, and then a range where the 
dependence of spark advance on flap setting is essentially negligible, and 
then a relatively rapid transition, and then, at the smallest flap 
openings, a range where MBT spark advance is again insensitive to flap 
setting. (A significant amount of data like that of FIG. 68 has been 
accumulated and the qualitative performance of the flap with respect to 
MBT spark setting is typical for a significant range of speeds and loads. 
Several points with respect to FIG. 68 are of interest. First, within 
certain rather well defined ranges, spark advance is not a sensitive 
function of flap setting. This is important, since it means that the spark 
setting will not be too sensitive to changes in flap setting. Another 
point, of course, is that the variable restriction flap does have a quite 
significant effect on flame speed. 
FIG. 69 plots the same experimental data points shown in FIG. 68 in a 
different way. In this run the indicated mean effective pressure of 
operation is plotted against flap opening. Several points with respect to 
FIG. 69 are important. First, it can be seen that the first effect of 
restricting the fluidic port is a small but definite and reproducible 
decrement in engine efficiency. Only when the flap restriction is 
significantly restricted is the efficiency of the engine improved with the 
variable restriction. It should be noted that an extrapolation of engine 
performance from small port restrictions would lead to the erroneous 
conclusion that additional engine turbulence infallibly hurt efficiency, 
while improvements in fuel economy could be obtained if the port 
restrictions were restricted more significantly. An additional point is 
commercially important. The engine output is not too sensitive to 
variations in flap setting. The flap setting is not such a critical 
parameter that it is difficult to control commercially. Relatively 
primitive flap actuation arrangements are therefore satisfactory, and the 
cost of the system is therefore much less than it would be if the engine 
performance was very sensitive to flap setting. 
FIGS. 70 to 75 show the sensitivity of the NO.sub.x output to variations of 
spark advance and variations of equivalence ratio. These curves should 
give a clear sense of the importance of proper equivalence ratio control 
and proper spark timing. FIG. 70 plots NO.sub.x with respect to variations 
in equivalence ratio for Blumberg's matrix point one. Lines are plotted 
for MBT spark timing, for MBT minus 5 spark timing, and additional points 
for MBT minus 7.5 degree spark timing are shown also. The penalty for 
enrichment of the mixture beyond the best economy mixture is substantial 
with respect to NO.sub.x. For the low load conditions of Blumberg's point 
one, there is also a slight but relatively insignificant penalty in being 
leaner than the optimal setting. 
FIG. 71 also shows the relationship between NO.sub.x, equivalence ratio, 
and spark advance for Blumberg's point one. The lines are relatively 
clearly labeled. The graph should make quite clear that there are 
substantial penalties to advancing the spark beyond the MBT spark setting, 
and significant advantages in retarding the spark from the MBT setting 
with respect to NO.sub.x. FIG. 72 is analogous to FIG. 71 but plots data 
for Blumberg's matrix point 4. For this higher load point, the penalty in 
nitric oxide of additional spark advance is small for mixtures leaner than 
the true best economy mixture. Also, it can be very clearly seen from FIG. 
72 that if the mixture is richer than the true best economy mixture, 
NO.sub.x levels are much increased and penalties from spark advance are 
increased also. FIG. 73 also deals with Blumberg matrix point 4 and is 
analogous to FIG. 70. Note that there is an extremely heavy penalty of 
mixture enrichment beyond the best economy mixture. FIGS. 74 and 75 plot 
NO.sub.x sensitivity to equivalence ratio and spark advance for Blumberg 
matrix point number 5 which is a higher load point. From these figures the 
very significant advantage of mixture enleanment should be clear, and the 
very significant NO.sub.x advantage of spark retard should also be clear. 
Specifically, for FIG. 75, note that with a mixture which is 3 percent 
leaner than the true best economy mixture (a mixture which involves less 
than a one percent fuel penalty) the very heavy penalty with respect to 
spark advance is very much softened from that which occurs with the richer 
mixtures. 
Those skilled in the internal combustion engine arts will recognize that 
the NO.sub.x sensitivity of the variable restriction engine to variations 
in equivalence ratio is significantly less than that shown by the 
three-way catalyst engine. Nonetheless, it is very clear that calibration 
of the engine is vitally important for low NO.sub.x to be achieved 
simultaneously with good fuel economy. Considerations of FIGS. 70 to 75 
should also make clear how heavy the penalties of cylinder-to-cylinder 
statistical variation can be, even if smooth engine operation were 
attainable with the cylinder-to-cylinder variations. It should also be 
noted that there are very substantial advantages in operating the engine 
slightly leaner than the true best economy mixture (or improving mixing so 
as to somewhat widen flame stability limits and lean out the true best 
economy mixture): As the mixture becomes more lean, the system is much 
more forgiving of spark advance variations and small percentage variations 
in equivalence ratio about the set point. With respect to the data 
involving the Blumberg correlation, it should be emphasized again that 
these data points were taken without the vortex mixer, and that 
substantial reductions in NO.sub.x output beyond those shown should be 
attainable with the addition of the vortex. Although a complete data set 
for the Blumberg points were not taken with the vortex, for Blumberg point 
5, which is responsible for a very significant fraction of the total 
NO.sub.x output for the correlation, the NO.sub.x output was reduced by 
more than a factor of ten with use of the vortex, and the fuel consumption 
was simultaneously improved. 
The structured flow and turbulence homogeneous charge engine of the present 
invention has the potential for NO.sub.x emissions very much below any 
proposed emission standard for NO.sub.x, with the hydrocarbon emissions in 
a calalytically controlled range. With respect to the catalytic control, 
it should be mentioned that the lean operation of the engine will 
eliminate the requirement for an air pump. Significant oxidation catalysis 
can be purchased for the manufacturing cost of an air pump with its 
auxiliary equipment. It should also be pointed out that the data shown 
here is probably not near the limit of the NO.sub.x control available with 
the fluidic port engine. With the use of exhaust gas recirculation, vortex 
mixing, or efficient fluidic ports, and some detail design with respect to 
combustion chamber shape, it is expected that the results can be 
significantly improved. This is important, since it is commercially useful 
to work with designs which produce emission levels significantly below the 
emission control targets, to compensate for the problems of unavoidable 
manufacturing variations. The very homogeneous charge engine of the 
present invention should permit NO.sub.x emissions to be held so low that 
NO.sub.x control is not a determining factor in engine design, so that 
engines can be built for best fuel economy and with excellent 
driveability. 
Because the engine of the present invention operates in a lean and 
homogeneous manner, its carbon monoxide emission rates are very low (so 
low that they could not be accurately measured with equipment available at 
the Internal Combustion Engine Research Laboratory of the University of 
Wisconsin). It should also be mentioned that the fuel economy predicted 
for the variable restriction port engine was 11.5 percent better than that 
of the baseline Ford engine. With more usual assumptions with respect to 
engine friction, the advantage of the variable restriction engine with 
respect to fuel economy would have been larger. The fuel economy penalty 
from the optimum optimorum for a 5 degree spark retard was approximately 
two percent, so that operation of the engine with a 5 degree spark retard 
would still have substantial fuel economy advantage over the baseline Ford 
engine. 
Other data with respect to the fluidic port variable restriction engine 
follows. FIG. 76 compared the NO.sub.x output at the lean limit (defined 
operationally as that equivalence ratio producing 6,000 parts per million 
methane equivalent) for the engine equipped with the flap and for the 
engine operated without the variable restriction. In both cases the engine 
was port injected. The lean limits were investigated over a substantial 
range of indicated mean effective pressures (engine loads). It can be 
clearly seen that the NO.sub.x output at the lean limit is drastically 
lower with the variable restriction employed than when the variable 
restriction is fully open. Usually this reduction in NO.sub.x output 
amounts to more than a factor of ten reduction. It should be emphasized 
that the NO.sub.x output levels would be lower yet if the vortex had been 
employed. FIG. 77 plots the indicated mean effective pressure and the 
indicated thermal efficiency measurements recorded for lean limit run 
described in FIG. 76. The advantage of the variable restriction flap can 
be clearly seen. 
FIG. 78 shows the effect of the flap at the very large opening of 0.400 
inches compared with the no flap case. A flap with an opening of 0.400 
inches under the 1200 r.p.m. engine operating conditions plotted here has 
an almost zero pressure drop across it, and therefore a relatively quite 
small velocity of flow across the port flow restriction. Nonetheless, it 
can be seen that the variable restriction systematically reduces the best 
torque spark advance, by changing structured flow patterns. 
FIG. 79 plots indicated specific HC versus equivalence ratio for the same 
data points as those plotted for FIG. 78, again where the flap engine was 
set up for 0.400 inches. For the relatively rich mixtures, the flap has 
only a small advantage, but for very lean mixtures the HC emission levels 
are very substantially less with the flap than they would be without it. 
FIG. 80 is again a plot of the same data points plotted for FIG. 78 and 
FIG. 79 and plots the indicated thermal efficiency of the engine versus 
its equivalence ratio. On the no flap run there is what appears to be a 
wild point which produces an anomalous hump in the curve, but the plot has 
been plotted to include it since the point was not shown to be wrong by 
actual checking. In all events, for the flap case, it can be seen that the 
indicated thermal efficiency of the engine (likewise the brake thermal 
efficiency) of the engine continues to increase as the mixture is enleaned 
all the way to the leanest point run. For the case of FIG. 80, it can be 
seen that the minimum NO.sub.x point, the leanest point, is also the best 
engine efficiency point. 
FIGS. 81, 82, 83, 42, and 43 plot various aspects of a Schweitzer curve 
(set fuel variable air flow spark advance optimal) run. FIG. 81 plots 
indicated thermal efficiency versus equivalence ratio. For this very low 
load, there is more scatter than would be desirable. However, the flap 
variable restriction points are generally significantly superior to the no 
flap points. FIG. 82 plots the same points in a manner which makes them 
look significantly more coherent, by plotting indicated specific fuel 
consumption in pounds of fuel per indicated horsepower hour versus 
equivalence ratio. Here, the superiority of the flap is somewhat clearer. 
FIG. 83 plots the same points plotting indicated specific hydrocarbon in 
grams HC per indicated horsepower hour versus equivalence ratio. It can be 
clearly seen that at relatively rich ratios there is no significant 
difference between the engine equipped with a flap and the engine with the 
flap restriction. However, in the very lean range the variable restriction 
flap is substantially superior to the conventional engine with respect to 
HC emissions. FIG. 42 plots MBT spark advance (a variable analogous to 
flame speed) versus equivalence ratio. It can be clearly seen that the 
flame speeds with the variable restriction very substantially faster than 
the flame speeds which occur without the restriction in the port. FIG. 43 
shows the superiority of the fluidic port variable restriction engine much 
more dramatically than FIGS. 81 to 83. In this relatively low load regime, 
the importance of in-cylinder mixing with respect to residual gases is so 
important that each and every data point plotted shows that the flap 
produces NO.sub.x emissions substantially below the NO.sub.x emissions 
which would occur without the flap. A consideration of the discussion with 
respect to FIGS. 3 to 6 should make this data quite reasonable: The data 
shown in FIG. 43 should clearly emphasize the importance of in-cylinder 
homogeneity for NO.sub.x control. 
FIG. 44 shows the MBT spark timings which correspond to the data points of 
FIG. 60. The increase in flame speeds with the flap restriction in the 
fluidic port are clearly shown. It is also clearly shown that the 
homogeneity of the mixture is not a significant determinant of flame 
speed, since the MBT spark timings of the port injected flap case and the 
vortex mixed flap case are not significantly different. This, of course, 
is consistent with the data and theory that flame speed is determined by 
in-cylinder turbulence levels. 
FIG. 45 shows peak pressure variation measured from oscilloscope 
photographs of the trace from the Kistler pressure probe. It can be seen 
that the variation in peak pressure with the flap is less than the 
variation in peak pressure without the flap. The data points on FIG. 45 
are also plotted in FIG. 60. FIG. 46 plots intake manifold vacuum in 
inches of mercury versus equivalence ratio. The most significant point 
with respect to FIG. 46 is that with a 0.300 inch flap opening at 1200 
r.p.m. the intake pressure drop across the flap restriction amounts to 
only about two inches of mercury. The flap has a significant combustion 
effect (shown dramatically in the NO.sub.x data plotted in FIG. 60) 
without having much of an effect on intake manifold vacuum. It should also 
be emphasized that FIG. 46 shows that even under the very leanest 
conditions in FIG. 60, further mixture enleanment was not limited by 
intake manifold vacuum, but was limited by flame stability. Skilled 
automotive engineers will recognize that a significant increase in torque 
for a set equivalence ratio is possible by changing intake manifold vacuum 
from four inches of mercury to roughly zero inches of mercury (something 
like a 40 percent increase). The very lean operating conditions for 
ultralow NO.sub.x emissions are possible under relatively quite high 
engine loads, although for absolute maximum power the fuel-air ratio must 
be richened to the maximum power ratio and under these (extremely 
infrequently encountered) conditions NO.sub.x output from the engine will 
be high. 
FIG. 47 plots indicated specific NO.sub.x versus equivalence ratio for a 
load condition higher than any which occurs in the Blumberg correlation. 
For this particular run, it can be seen that the flap is actually worse 
with respect to NO.sub.x at the mixtures which correspond roughly to the 
maximum output level, but that the variable restriction becomes 
advantageous under the (fuel economy optimal) leaner ratios. It should be 
noted that the data of FIG. 47 was run with the engine port injected and 
that the minimum NO.sub.x values could have been reduced substantially 
with addition of the vortex. 
FIGS. 48 and 49 compare the inventor's fluidic port data to data published 
by Peters and Quader of General Motors Laboratories on a stratified charge 
engine ("Wetting the Appetite of Spark Ignition Engines for Lean 
Combustion" S.A.E. paper #780234, July 1978). 
FIG. 48 plots indicated specific fuel consumption in micrograms per joule 
versus equivalence ratio. The fluidic port data with respect to fuel 
economy is substantially as good as that of the best injection timing 
(BIT) data in the figure for the engine stratification scheme. FIG. 49 
plots the inventor's fluidic port variable restriction data versus Peters' 
and Quader's plot. It can be seen that the fluidic port plot is very 
similar to that of Peters' and Quader's premixed plot except that the lean 
misfire limit for Peters' and Quader's data is very much richer than the 
lean misfire limit for the fluidic port engine. Because the plot in FIG. 
49 is not a logarithmic plot on NO.sub.x, the drastic reduction in 
NO.sub.x for the fluidic port leaner than 0.62 stoichiometric cannot be 
plotted. However, it is important to compare the NO.sub.x performance of 
the fluidic port engine at its best economy (TBEM or true best economy 
mixture) with that for the Peters and Quader engine. For like equivalence 
ratios the Peters and Quader engine, because it involves charge 
heterogeneity, has much higher NO.sub. x output levels. Under conditions 
leaner than 0.62 equivalence ratio, when the NO.sub.x output with the 
fluidic port engine is almost zero, the NO.sub.x output of the Peters and 
Quader engine is still substantial. In addition, the lean misfire limit 
for the fluidic port engine is as lean as the lean misfire limit for the 
Peters and Quader (wet injection) stratified charge engine. It is worth 
pointing out that the Peters and Quader work has been considered important 
enough to be significantly noticed in the automotive engineering press, 
and that in the writeup of the Peters and Quader paper, no clear 
understanding of the stratification process in the Peters and Quader 
engine (which involves a structured flow process) existed. 
DISCUSSION OF PHRASES IN THE CLAIMS 
In a case so long and so complicated as this one, it is well to have a 
discussion of some of the terms used in the language of the claims in a 
form compact enough for ready reference. Since the present case involves 
interrelated concepts either novel or quite infrequent in automotive 
engineering, precise definition of the terms used in the claim language is 
also somewhat more important than it might be otherwise. 
The present invention intimately involves the use of controlled, 
structured, turbulent flows to produce intimate mixing and adequate flame 
speeds for the very dilute mixtures required for very low NO.sub.x 
emissions. Producing these controlled, structured, turbulent flows 
involves attention to details of flow geometry not previously appreciated 
in automotive engineering. The terms involved are, however, well 
understood by men skilled in the art of fluidics and references to this 
fluidic field have already been given. A high speed flow near a reasonably 
well shaped passage wall or surface will tend to attach to that surface to 
form a wall attached "Coanda" stream, and this stream will tend to spread 
and dissipate its kinetic energy into random turbulence much more slowly 
than would an unattached, free jet stream. It is this Coanda wall 
attachment effect which permits a significant fraction of the flow energy 
past a variable port restriction to be delivered into an engine combustion 
chamber in sufficiently coherent form to produce a useful structured flow 
in the cylinder. Some of the fluid mechanics of the wall attachment effect 
has been discussed in this case, and much more detailed descriptions exist 
in the fluidic literature. If the flow energy past the variable 
restriction is to be preserved for delivery into the engine cylinder, it 
is important that the port be "shaped to preserve the Coanda flow." A 
number of rules, sufficient to guide one skilled in fluidics or fluid 
mechanics in the shaping of a fluidically efficient port, have been given 
in the current case, and an important negative rule, the avoidance of 
step-up, has also been discussed. Since an infinite number of small 
variations in port shape is possible, a complete enumeration of all port 
shapes, "shaped to preserve the Coanda flow," is not possible, but one 
skilled in fluidics should be able to in most cases tell whether a port is 
shaped to preserve the flow or not. In all cases, the ability of the port 
to preserve the energy in the Coanda flow can be determined by simple 
steady state flow experiments which are very inexpensive in terms of the 
issues involved. 
The concept of a structured turbulent flow is central to the present 
invention. In a structured turbulent flow, most of the kinetic energy in 
the flow exists in the form of a relatively well defined pattern, or 
hydrodynamic dance, and random turbulence is superimposed on this 
structured flow pattern. Much effort in the current case has gone towards 
explaining the concept of structured turbulent flow and showing its great 
relevance to internal combustion engine mixing. In the present invention, 
very rapid mixing is attained by having the flow energy from the variable 
port restriction produce a structured turbulent flow in the cylinder where 
any initial concentration field is stretched, distorted, and spread so 
that even if charge stratification existed early in the mixing process, 
the flow would distribute the rich and lean elements widely within the 
volume to be mixed so that the mean distance across which random turbulent 
molecular diffusion would have to occur in order to complete the mixing 
process would be relatively short. An important example of this sort of a 
rapidly mixing turbulent flow structure is the irrotational flow vortex, 
although a great many other efficient mixing flow structures also exist. 
An example has also been given of a turbulent structured flow relatively 
very ineffective for mixing in the form of the rigid body rotation vortex. 
Flow structures characterized by high velocity gradients tend to stretch, 
distort, and spread the mixants in a desirable way. 
The term turbulence, in the sense in which it is used with response to 
structured turbulent flows and in the sense in which it is used in the 
claims, is the random fluctuating motion which may be distinguished from 
the structured flow motion. The concept of turbulence as it is used here 
is rather well exemplified by the analysis of David Lancaster's data shown 
in FIGS. 18A, 18B, and 18C of the present case. 
The present invention is intimately concerned with questions of mixing 
which are fundamentally statistical issues. Terms such as "completeness of 
mixing," "very tight statistics," and "tight air-fuel-residual microscale 
mixing statistics" occur in the claims. A distribution is held to be 
"tight" when the standard deviation of individual elements from the mean 
is small. As the graphical illustration of FIG. 6 showed, tight air-fuel 
mixing statistics permit adequate combustion with much leaner ratios than 
would be permissible with worse statistics. It should be pointed out that 
the best combustion results under the present invention require tight 
cylinder-to-cylinder air-fuel-residual statistics, tight statistics cycle 
to cycle for each cylinder, and then mixing inside the cylinder and 
combustion chamber so that the microscale fuel-air residual mixing 
statistics are also tight. Means to achieve these goals using structured 
turbulent flows have been disclosed in the current case. 
The excellent mixing and flow control disclosed in present case will not 
produce low nitric oxide emissions unless the engine is operating with a 
dilute fuel-air-residual mixture. For example, the data of FIG. 60 show 
that unless the mixture is quite dilute, there is essentially no NO.sub.x 
advantage to the controlled structured turbulent flow system, but that 
with sufficiently dilute mixtures, very dramatic reductions in nitric 
oxide output are available with the improved mixing and controlled 
turbulence. The definition of the properly dilute mixture is made 
complicated by the fact, illustrated in FIG. 3 that the mixture may be 
made dilute by the addition of air in excess of the stoichiometric 
proportion, or by exhaust gas recirculation products of previous 
combustion, or by some combination of exhaust recirculation and excess 
air. In the claims, dilute combustion means combustion with a 
fuel-air-residual gas mixture leaner than the stoichiometric mixture and 
having a high enough concentration of diluents (some combination of excess 
air and products of previous combustion) so as to have low NO.sub.x output 
at best power spark timing. In some of the claims, an attempt is made to 
define the required level of mixture dilution in a determinant way, and 
this definition necessarily involves an experimental determination of the 
misfire limit dilution corresponding to the engine speed-load map at a 
number of points. This experimental determination of misfire limits is 
straightforward and relatively inexpensive and forms an unambiguous test 
as to the level of mixture dilution required.