Torch ignition apparatus and method

A torch ignition apparatus and method, the apparatus including a secondary, torch ignition piston interconnected to the primary piston and operable to produce through compression ignition of a carbureted fuel/air mixture a jet of hot gases capable of igniting a fuel and air mixture compressed in a primary combustion chamber by the primary piston. The apparatus also includes a secondary fuel system for supplementing a low ratio fuel/air mixture in the combustion chamber and a novel hoop valve and exhaust port system for providing a unique, highly efficient flow-through system for exchanging gases in the combustion chamber.

BACKGROUND 
1. Field of the Invention 
The present invention relates to ignition systems for internal combustion 
engines and, more particularly, to a novel intake, exhaust and torch 
ignition apparatus for a combustion chamber of a cylinder of an internal 
combustion engine. 
2. The Prior Art 
Historically, internal combustion engines incorporate either a compression 
ignition system (diesel-type), a spark ignition system or variations of 
these systems to ignite the fuel/air mixture in the combustion chamber. 
The spark-ignition engine utilizes carburetion (a premixing of the fuel, 
usually gasoline, with air) prior to directing the fuel/air mixture into 
the combustion chamber. The conventional carburetor is a relatively 
inefficient device in that much of the gasoline is not completely 
vaporized but remains in droplet form as it enters the combustion chamber 
with the result that some of these droplets are discharged into the 
exhaust as unburned hydrocarbons. Additionally, the combustion temperature 
from the spark-initiated ignition is relatively low with the result that 
there is incomplete combustion so that additional quantities of gasoline 
are discharged into the atmosphere as unburned or incompletely burned 
hydrocarbons and carbon monoxide. While strenuous efforts have been 
directed toward the reduction of these pollutants by various devices 
including the injection of supplemental air into the exhaust system to 
continue the combustion process, catalytic convertors, and the like, only 
modest results have been obtained in spite of the high cost of the 
apparatus. 
Attempts to increase the efficiency of the internal combustion engine and 
also the horsepower available have succeeded dramatically by using higher 
compression ratios on the order of about 10:1 or even 11:1. However, 
gasoline frequently detonates spontaneously at these higher compression 
ratios with an adverse effect on engine performance and life. The addition 
to gasoline of detonation inhibitors such as tetraethyl lead, 
tricresolphosphate, and the like, inhibits detonation thereby 
accommodating high compression ratios. However, since lead is a 
significant pollutant, numerous governmental regulations have been adopted 
against its usage. Accordingly, the current gasoline or spark-initiated 
engines are directed toward using unleaded gasoline with the result that 
they cannot be operated at compression ratios much greater than about 9:1 
without fear of detonation. 
In summary, the current spark-initiated, internal combustion engines have a 
relatively low efficiency for the following reasons: 
(1) Lower compression ratios, 
(2) Poor combustion as a result of lower compression ratios and consequent 
low maximum operating temperature, 
(3) Poor combustion as a result of poor carburetion, and 
(4) Governmentally-mandated anti-pollution devices which have uniformly 
reduced engine efficiency. 
The compression ignition or diesel-type engine is inherently more efficient 
than the spark-initiated ignition engine since (1) there is more energy 
per weight in diesel fuel than gasoline and (2) the engine operates at a 
higher combustion temperature and compression ratio. Customarily, the 
compression ratio for a relatively small engine, such as in an automobile, 
is as high as 23:1 in order for the temperature to be high enough to 
ignite the injected fuel. The higher combustion temperatures also mean a 
more thorough combustion of the hydrocarbons and carbon monoxide with a 
corresponding increased energy release and consequent greater mechanical 
energy produced for the same volume of fuel consumed. Admittedly, 
compression-ignition engines characteristically exhibit a lower horsepower 
per pound of fuel consumed than spark-ignition engines. This lower 
horsepower generally results from the high peak pressures (which, in turn, 
require a heavier engine structure) and from other problems which exceed 
the savings realized from the more thorough combustion. 
Ignition in a compression-ignition engine results when an air induction 
charge is compressed in the combustion chamber to a relatively high 
pressure and a correspondingly high temperature above the ignition 
temperature of the fuel so that subsequently injected fuel ignites as it 
is injected. The fuel must be injected rapidly, or almost all at once, so 
that there will be at least some of the injected fuel that is an 
appropriate mixture of fuel and air for immediate autoignition. However, 
other parts of the injected mixture will be typically too rich for 
ignition while some parts will be too lean, although injection almost all 
at once insures that at least some of the fuel and air will be at the 
appropriate mixture for instantaneous ignition. 
On the other hand, if the fuel is injected slowly, ignition will not start 
instantaneously because there will not always be a correct mixture for 
ignition at the instant of injection. After enough fuel has finally been 
injected so that some of it is at the correct mixture, it will ignite. 
However, by the time ignition occurs, a considerable quantity of fuel that 
had previously been injected will be relatively close to the correct 
ignition mixture so that it will ignite from the mixture that actually was 
first ignited, although later injected. Upon this occurance, the large 
quantity of previously injected fuel ignites extremely rapidly with a 
consequent very rapid and abnormally high rise in the cylinder pressure 
and temperature. This phenomena is essentially equivalent to a secondary 
detonation of "carbureted fuel." This is not desirable and can result in 
piston and cylinder damage. As a result, a small to medium-size, 
diesel-fueled, compression-ignition engine requires that all of the fuel 
that is supplied to a cylinder in a particular stroke or revolution must 
be injected at once in order to prevent the subsequent detonation caused 
by the combination of (a) carburetion and (b) delayed ignition that occurs 
when the fuel is injected slowly. 
Although it is necessary for a rapid and full injection of fuel to prevent 
abnormal and unwanted detonation of "carbureted fuel", this very process 
causes most of the problems in small automobile-sized diesel engines. When 
all of the fuel is injected, the subsequent pressure and temperature rise 
is virtually uncontrolled and both rise to very high values early in the 
power part of the stroke. This pressure rise is so rapid that the peak 
pressure occurs before the sine of the power angle is large enough to 
produce very much torque. For example, at top dead center (TDC) with the 
power angle at 0 degrees, any amount of pressure would produce zero torque 
since the sine of 0 degree is zero. Furthermore, if the peak pressure 
occurs at a power angle of about 10 or 15 degrees, the sine of the angle 
is still modestly low so that not much torque results. However, the 
structure of the engine must be sufficiently strong so as to be able to 
withstand these relatively high pressures. Engine strength is usually 
defined by engine weight so that the resulting heavy piston and connecting 
rod parts restrict the maximum speed of the engine and make it relatively 
sluggish so that it cannot increase in speed rapidly or operate very fast. 
A second problem arises from the requirement that the diesel fuel be 
injected all at once. When autoignition occurs and thus initiates the 
overall combustion process in the cylinder, secondary turbulence produced 
by the initial combustion causes the remainder of the fuel to mix with air 
in the cylinder so that the combustion process can continue. However, this 
process is not very efficient and much of the fuel is never suitably mixed 
with air. The unmixed fuel is turned to carbon particles by the high 
temperatures and is discharged into the exhaust as soot. Exhaust soot is 
the primary pollutant in a small diesel engine, and it also infiltrates 
into the oil system necessitating relatively frequent oil changes. 
Although carburetion would be an excellent solution to the soot problem of 
the diesel engine so that essentially all of the fuel could be burned, the 
very fact that carbureted diesel fuel burns so rapidly--approaching or 
encompassing detonation speeds and pressures with attendant problems 
described above--makes full carburetion an almost impossible, idealistic 
goal. 
A third problem arises from the high temperature that results from having 
to inject all of the fuel at once. The high temperatures that occur in a 
diesel engine are so high that a large percentage of the nitrogen combines 
with oxygen and is discharged into the air as oxides of nitrogen. 
As noted before, combustion pressure is almost completely uncontrollable in 
spark and compression ignition engines. About all that can be controlled 
in these engines is the maximum pressure. Subsequent power angle pressures 
largely follow the formula PV/T=C, and are definitely not optimum for 
maximizing efficiency and horsepower while minimizing pollutants. 
While the fuel-air charge in most spark ignition engines at the time of 
combustion is essentially a homogenous and vaporized mixture of fuel and 
air; if the distribution of fuel is not uniform within the chamber, zones 
of varying air/fuel ratios will be present. Such a mixture is termed a 
stratified charge. For example, in a stratified charge, the air/fuel ratio 
at one point in the chamber might be 16:1 while only air might exist at 
another point in the chamber. The purposes of the stratified charge, spark 
ignition engine are to: (1) permit use of a leaner mixture than could 
ordinarily be used in producing ignition successfully; and (2) avoid knock 
with the result that either high compression ratios or low-grade fuels (or 
both) can be used. A stratified charge can theoretically eliminate knock 
because the end gas need not be a combustible mixture. The residence 
(heating time) of the fuel is also short because injection begins late in 
the compression stroke. 
The stratified-charge principle is one technique used for obtaining high 
compression (expansion) ratios in combination with spark ignition. 
However, it is difficult to initate combustion in lean mixtures by a spark 
discharge and, since mixtures are never perfectly homogeneous, several 
regions might necessarily have to be ignited to assure continued flame 
propogation. Additionally, the propogation of the flame becomes 
increasingly slower as the ratio of fuel to air is reduced, until it 
virtually ceases at values of approximately 0.025 (fuel-air ratio). 
Accordingly, the release of energy arising from, say, two ignition points 
would be extremely slow. One solution to this problem is offered by the 
use of a dual-fuel diesel engine. In this engine, a homogenous and lean 
mixture of gas and air is compressed to a high pressure and temperature 
and thereafter ignited by injecting a small pilot charge of fuel oil. The 
small spray of oil establishes a large number of ignition points, not on 
the edges of the chamber as with sparkplugs, but throughout the entire 
gas-air mixture. The mixture ratio in the vicinity of the oil droplets 
will be enriched and combustion will start smoothly and rapidly. A number 
of flame fronts will thereby be established, although as each flame 
penetrates into the gas-air mixture, its progress will become slower. In 
fact, if the air-gas ratio exceeds about 40 to 1, the flame may be 
extinguished in part, as evidenced by unburned fuel in the exhaust. 
Therefore, it is interesting to note that in the dual-fuel engine, 
combustion starts in similar fashion to a compression ignition engine and 
combustion continues by flame propogation, in a similar fashion to a spark 
ignition engine. The advantage of a dual-fuel engine is that it will 
exceed the performance of a straight diesel engine at full load, since 
vaporized gasoline or gas is present in all parts of the chamber and 
therefore, more air can be burned. 
The principle of one experimental engine can be visualized by assuming a 
circular motion of the air in the cylinder on the compression stroke. At 
about 50.degree. before top dead center, let a nozzle start to inject fuel 
tangentially into the air stream and continue the injection for, say 
50.degree. of crank movement (at full load). Meanwhile, in a position 
downstream from the nozzle, a sparkplug is located, and, after injection 
begins, a spark occurs (say 30.degree. before TDC) when initial fuel-air 
mixture is swept by turbulence (air swirl) past the sparkplug. Here the 
flame will be initiated and propogated, mainly in a direction opposite to 
the swirl with establishment of a burning zone. The liquid fuel leaving 
the nozzle will vaporize, mix with air, and then burn, thus establishing 
the lower boundary of the zone. The products of combustion will be carried 
out of the burning zone and swept around the combustion chamber. In this 
manner, the fuel/air mixture will be burned directly after the fuel enters 
the chamber, and without waiting for a flame to travel to the mixture 
position in the combustion chamber. 
Conventionally, the develoment of an injection-type, stratified-charge 
engine introduced a new problem of coordinating the injection of fuel with 
the design and turbulence of the combustion chamber. This coordination has 
not only been difficult to achieve, but it has been indicated to be rather 
improbable that one design could successfully operate over wide limits of 
loads and speeds utilizing the present systems for ignition. In 
particular, the nozzle must not only give good atomization but also 
selective distribution: a local combustion area must originate and develop 
from the spark plug location. Moreover, the nozzle and turbulence must, in 
some manner, follow the principles dictated by volume distribution, if the 
pressure rises are to occur at the most advantageous crank angles. It has, 
therefore, been indicated to be quite improbable that the process will 
give, in itself, such optimum pressure-like characteristics. The primary 
advantage to the injection-type, spark ignition engine is that it can 
handle a number of fuels (fuel oil or gasoline) while other engines are 
more particular since knock, a destructive process, can intervene. 
A third type of engine is the torch-ignition engine. It utilizes neither a 
spark nor high compression to initiate combustion. Ignition is produced at 
the proper part of the stroke by extremely hot gases in the form of a 
torch or searing flame. Inasmuch as the torch technology is not subject to 
the inhibiting deficiencies of either the spark or compression-type 
engines--as will be shown--all of the weaknesses of these other two kinds 
of engines can either be ameliorated or wholly eliminated, thus increasing 
efficiency and horsepower while decreasing the presence of pollutants. For 
example, the optimum compression ratio for a small automobile engine is 
about 16:1. At this compression ratio, efficiency and horsepower can be 
maximized and the creation of pollutants can be minimized. While a torch 
can ignite a fuel at a compression ratio of 16:1, a carbureted gasoline 
spark-initiated engine will detonate at that ratio and a small compression 
ignition or diesel-fueled engine would not ignite at all at that ratio 
unless very hot. Inasmuch as efficiency increases rapidly up to a 
compression ratio of about 16:1, controlled fuel injection with a torch 
ignition system can always be operated at the most efficient compression 
ratio because ignition is always assured and timing of ignition is 
completely controllable. 
Torque, and hence horsepower, can also be maximized at optimum power angles 
while concurrently limiting extreme maximum pressures and temperatures. 
This is accomplished by injecting fuel at the optimum angle near TDC, and 
regulating or controlling the injection rate so as to predetermine the 
resulting pressure rise from combustion at the desired optimum value for 
each arc of the power angle--the angle of the crankshaft between TDC and 
commencement of exhaust. 
In summary, each prior art system is directed toward providing an internal 
combustion engine that is more economical in its consumption of fuel 
and/or more efficient in the fuel combustion. However, none of the known 
devices relate to an internal combustion engine apparatus and method 
whereby a lean fuel/air mixture is contained within the combustion chamber 
while a flame front is created within the combustion chamber by injecting 
high temperature gasses and additional fuel into the combustion chamber. 
In view of the foregoing, it would be a significant advancement in the art 
to provide a novel internal combustion engine apparatus and method whereby 
a flame front is created within the compressed air or lean fuel/air 
mixture in the combustion chamber of an internal combustion engine, the 
flame front being created by injecting high temperature gasses and 
additional fuel into the combustion chamber. It would also be an 
advancement in the art to provide a novel internal combustion engine 
apparatus and method whereby the fuel/air mixture is introduced into the 
combustion chamber in a swirling motion to thereby continuously feed the 
fuel/air mixture through the flame front during the combustion cycle. It 
would also be an advancement in the art to provide a novel intake and 
exhaust apparatus and method thereby readily adapting the apparatus and 
method of this invention to an internal combustion engine. Such a novel 
internal combustion engine apparatus and method is disclosed and claimed 
herein. 
BRIEF SUMMARY AND OBJECTS OF THE INVENTION 
The present invention relates to a novel apparatus and method for 
initiating combustion in a combustion chamber of an internal combustion 
engine. Combustion is initiated by establishing a flame front within the 
combustion chamber, the flame front being formed by injecting a high 
temperature gas and additional fuel into the combustion chamber. The 
additional fuel supplements the lean fuel/air mixture or air-only charge 
in the combustion chamber while the high temperature gasses raise the 
temperature of the resulting fuel/air mixture above the ignition 
temperature of the mixture. Novel intake and exhaust systems are also 
included for the purpose of more efficiently exchanging the gasses within 
the combustion chamber. 
It is, therefore, a primary object of this invention to provide 
improvements in internal combustion engines. 
Another object of this invention is to provide an improved apparatus for 
initiating combustion of a fuel/air mixture in a combustion chamber of an 
internal combustion engine by injecting high temperature gases--or 
torch--from a source independent of the fuel mixture in the main cylinder 
and establishing a flame front in the combustion chamber. 
Another object of this invention is to provide an improved method for 
initiating combustion of a fuel/air mixture in the combustion chamber of 
an internal combustion engine. 
Another object of this invention is to provide improvements in the intake 
and exhaust systems for a cylinder of an internal combustion engine. 
Another object of this invention is to provide an improved internal 
combustion engine apparatus wherein a secondary piston is utilized to 
ignite an ignition fuel/air mixture thereby providing high temperature 
gasses for establishment of a flame front within the combustion chamber of 
the internal combustion engine. 
These and other objects and features of the present invention will become 
more fully apparent from the following description and appended claims 
taken in conjunction with the accompanying drawing.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
The invention is best understood by reference to the drawing wherein like 
parts are designated with like numerals throughout. 
General Discussion 
Self-ignition or autoignition is a spontaneous chemical reaction that 
occurs when a mixture of fuel and oxygen spontaneously reacts without the 
necessity of a flame or spark to initiate combustion. When this occurs, 
the pressure and temperature abruptly increase because of the sudden 
release of chemical energy. The factors that control autoignition include 
(1) temperature, since if the temperature is high, the molecular energy is 
high and therefore, high energy molecular collisions can cause the 
formation of new molecules with a corresponding release of energy; (2) 
density, since if the density is high the number of collisions is many and 
therefore, the number of new molecules formed by collisions is great; and 
(3) reaction rate, since the rate of the reaction is also controlled by 
the relative numbers of reacting molecules (as represented by the air/fuel 
ratio) as well as by the presence of inert moleculs (such as nitrogen in 
the air) that influence the reactive molecular collisions. 
It is difficult to separate the effects of the various factors of 
autoignition. For example, if a homogenous fuel/air mixture were rapidly 
compressed and held at the high temperature and pressure achieved by the 
compression, the mixture may slowly cool without autoignition although an 
analysis of the mixture would undoubtedly show some signs of oxidation. 
However, if the compression is raised sufficiently, a state will finally 
be reached where self-ignition will occur. But even after this state is 
attained, an induction period is present and results in an ignition delay 
before the reaction becomes explosive. Accordingly, it is postulated that 
certain preflame conditions must occur during the induction period to 
condition the mixture for self-reaction. Although the exact mechanism of 
formation is unknown, it is believed that some intermediate product of 
combustion appears in the induction period and serves to catalyze the 
entire reaction to explosive speeds. When the mixture is compressed to 
higher temperatures than before, it is found that the ignition delay 
period is shortened considerably. This appears reasonable since, at the 
new state, the molecular activity is greater than before. Accordingly, it 
appears that autoignition of a perfectly homogenous mixture of gasses is 
controlled by several factors: (a) temperature, (b) density, (c) time, 
that is, the induction period, (d) composition including the fuel/oxygen 
ratio and the presence of inert gasses or any other substance that affects 
the chemical reaction, and (e) turbulence since if the mixture is not 
homogenous, a mixing factor is involved. 
The term "ignition delay" is assigned to the time consumed by both the 
physical and chemical delays in the ignition process. For light fuels, the 
physical delay is small, while for heavy, viscous fuels, the physical 
delay may be the controlling factor. The physical delay is greatly 
reduced, therefore, by using high injection pressures and high turbulence 
or other techniques to facilitate breakup of the fuel jet. In most 
compression-ignition engines, the ignition delay is shorter than the 
duration of injection so that the combustion period can be considered to 
be divided into four stages: (1) ignition delay; (2) rapid pressure rise; 
(3) constant pressure rise; and (4) burning on the expansion stroke. The 
rapid pressure rise occurs because of the myriad ignition points and the 
accumulation of fuel during the delay period and is uncontrollable once 
started. Following this stage, in large diesel engines the final portions 
of the fuel are injected into the flame and consequently combustion of 
this portion is somewhat regulated by the injection rate. However, in 
medium (truck) and small automobile-sized engines, injection must be all 
at once in order to prevent detonation as set forth hereinbefore. Since 
the process is far from homogenous, combustion continues when the 
expansion stroke is well underway so that this continued burning can be 
referred to as the fourth stage of combustion. 
It should be noted in the spark-ignition engine, ignition occurs at a 
single point, with consequent slow rise in pressure as compared to the 
compression-ignition engine wherein ignition occurs at many points with a 
consequent rapid rise in pressure. For this reason, the spark is initiated 
earlier, depending upon engine speed, on the compression stroke (say 
30.degree. before TDC) than does injection in a similar, but compression 
ignition engine (say 15.degree. before TDC). Also, in the spark ignition 
engine, the flame speed is primarily controlled by turbulence created 
before the start of combustion. The turbulence created before combustion 
starts is referred to as primary turbulence. In the compression-ignition 
engine primary turbulence assists in breaking up the fuel jet and 
continues in intermixing the burned and unburned portions of the mixture. 
Accordingly, most compression-ignition engines that depend on primary 
turbulence are limited to low rotating speeds as compared to the 
spark-ignition engine because of the heterogenous mixture in the 
combustion chamber. Those mixtures that are not correct for combustion 
will not ignite until they are correct. 
Maximum power output from the internal combustion engine is obtained when 
all of the oxygen in the cylinder is effectively consumed, thus releasing 
the chemical energy in the fuel. Accordingly, there will be an increase in 
power output with increase in fuel until a point is reached where all of 
the oxygen in the cylinder is effectively utilized for combustion. 
Although fuel flow can be increased while air flow is fixed by the design 
and displacement of the engine, it is, therefore, the air and not the fuel 
that imposes a limit to the power output. When it is realized (a) that the 
fuel and air are imperfectly mixed, (b) that the fuel may not be 
completely vaporized, and (c) that the clearance space is partially filled 
with exhaust products which dilute the concentration of fresh charge, it 
becomes evident that an excess of fuel, relative to the chemically correct 
mixture, must be employed to obtain maximum power. 
On the other hand, the maximum economy of the engine, as determined by the 
specific fuel consumption, should be obtained when the release of chemical 
energy, per unit of fuel, is a maximum. This condition is secured when the 
fuel is completely burned, and therefore an excess of oxygen must be 
supplied. Since both lean and rich mixtures burn slowly, the piston will 
descend on the power stroke while energy is being liberated. Also, the 
energy released is used, in part, to overcome friction which is relatively 
constant. Because of these undesirable factors, the mixture for maximum 
economy of the engine will, therefore, not depart far from the chemically 
correct mixture. In summary, maximum power occurs when all of the air is 
burned with a rich air-fuel ratio, while maximum economy is obtained when 
a lean mixture is employed. The difference in ratios between these two 
performance points is controlled by the design of the engine. 
In the torch ignition engine--using a two-cycle engine for ease of 
illustration--a charge of air is inducted into the cylinder through ports 
adjacent the top of the cylinder from a source of higher pressure such as 
a supercharger. As soon as the piston-operated exhaust ports are closed at 
the bottom of the cylinder, a partial charge of pure fuel is injected into 
a rapidly swirling air stream at the top of the cylinder and is thus 
carbureted within the cylinder. The air continues to flow in at the top 
until supercharged to the desired value and then the valve at the top 
closes and compression begins. 
At the proper part of the compression cycle near TDC, the remaining portion 
of fuel charge is injected at a predetermined and controlled rate. 
Concurrent with injection of the fuel at this time, an ignition torch is 
introduced along with the fuel injection spray and in a direction that 
enhances the swirling action of the charge of carbureted air and partial 
fuel. The torch instantly ignites the combined second fuel charge as it 
mixes with the first fuel/air mixture. The first fuel/air mixture is set 
as rich as possible short of what is required for combustion (which would 
produce detonation) and the second fuel charge provides the remaining 
amount of fuel needed for optimal combustion. If necessary, the torch can 
continue to be introduced during the time the second fuel is injected in 
order to assist the establishment of a stationary flame front in 
juxtaposition with the fuel injection nozzle. As the fuel is injected, a 
flame front is initiated and continues to combust with new fuel/air 
mixture from the incoming swirl while discharging combusted products 
downstream. In this manner, all of the fuel can be completely combusted 
while producing a pressure rise within the cylinder that is optimum for 
each sweeping degree of arc of the power angle. 
The combination of (a) relatively high temperature, (b) second fuel 
injection, (c) torch ignition, and (d) a rapidly swirling, 
partially-carbureted first fuel/air mixture insures that substantially all 
carbon monoxide and hydrocarbon pollutants are burned or never formed as 
such. Injection of the secondary fuel is regulated at a rate which will 
produce the desired pressure in the cylinder at all times. If the pressure 
is too high for part of the cycle, less fuel is injected during that part 
of the cycle, but if the pressure is too low, more fuel is injected. The 
fuel is thus injected at a rate that limits maximum pressure and 
temperature to modest values that are highest when the power angle will 
produce the greatest torque. Since maximum pressure is not excessively 
high, engine parts can be relatively light, such as in a spark-ignition 
engine, rather than extra strong and heavy as in a compression-ignition 
engine. 
Since maximum pressure is not high, maximum temperature is likewise not 
excessively high. Without high temperatures, the production of oxides of 
nitrogen is low, similar to a spark-ignition engine. The same conditions 
that ensure complete combustion of carbon monoxide and hydrocarbons also 
ensures complete combustion of the second fuel injection to avoid the 
production of soot. The rapidly swirling, lean carbureted first fuel/air 
mixture is quite easily mixed with the relatively slowly injected second 
fuel injection so that all fuel combines with oxygen and is combusted. In 
summary, engine efficiency and horsepower are fully optimized by (a) using 
an optimum compression ratio and (b) combusting all of the injected fuel. 
Various valve systems are utilized to induct air and expel exhaust over the 
desired speed range of an internal combustion engine. In a four-stroke 
cycle, the intake valve starts to open before TDC in order that the valve 
will be appreciably open near the start of the intake stroke. The intake 
valve remains open during the early part of the compression stroke to 
increase the charge of the cylinder at high speed and, also, to reduce the 
compression ratio at low speeds as a means for avoiding knock in the 
spark-ignition engine. 
In similar fashion, the exhaust valve opens before the power stroke is 
completed, and closes after the intake process has begun. The early 
opening of the exhaust valve insures that the pressure in the cylinder 
will approach atmospheric before the piston begins the exhaust stroke. By 
this timing, part of the expansion energy of the gasses on the power 
stroke is lost, but this loss is more than compensated by the decreased 
amount of work necessary for the engine during the exhaust stroke. The 
gain, therefore, is at wide-open throttle and full-speed operation while 
at slower speeds, early opening of the exhaust is decidedly a loss. An 
alternative course would be to increase the size of the exhaust valve and 
port, but this remedy has, historically, been considered impossible for 
lack of room, and is considered undesirable since the larger valve would 
run hotter. 
It should also be noted that both the intake and the exhaust valves may be 
open at the end of the exhaust stroke, which is also the beginning of the 
intake stroke and, therefore, the valves may be overlapped. The degree of 
overlapping is increased when maximum power at high speed is what is 
desired for then the scavenging process is encouraged. On the other hand, 
overlapped valves are not conducive to good part-load operation. In 
addition, some of the incoming carbureted mixture travels directly from 
the intake valve to the juxtaposed exhaust valve. The percentae of loss 
increases as the valve open overlapping time period increases. This 
unburned mixture comprises part of the pollutants in the exhaust that is 
the subject of much governmental concern. 
The most common two-stroke cycle engine controls the intake process 
entirely by piston movement convering or uncovering the intake and exhaust 
ports. In some engines, a scavenging blower completes the exhaust process 
by supplying fresh air to push the exhaust gas out of the cylinder so that 
some air is wasted in this process (but not fuel). 
Turbulence induced on the charging stroke and on the compression stroke, 
and also by the action of an injection spray itself, is referred to as 
primary turbulence. Once combustion has been initiated, the turbulence may 
be increased by the burning or exploding nature of the process, and this 
is referred to as secondary turbulence. Primary turbulence is most readily 
induced and, also, scavenging in a two-stroke-cycle engine is most 
efficient when flow is in one direction, uniflow. One method for obtaining 
uniflow is to locate the inlet ports in the cylinder and the outlet ports 
in the opposite end in the head. Another method is to use opposed pistons 
with the upper piston uncovering the inlet ports and the lower piston 
uncovering the exhaust ports. Additionally, all high-efficiency 
two-stroke-cycle engines employ scavening blowers to assist the exhaust 
and intake processes. The pressure in the intake manifold is maintained at 
a predetermined pressure and air cannot enter the cylinder until the 
scavenging pressure is less than the intake pressure and this occurs when 
the exhaust pressure falls below that of the scavening pressure. 
The present invention relates to a completely new type of internal 
combustion engine, the true torch ignition engine. Ignition is not 
obtained by a spark and is not initiated by high compression within the 
main combustion chamber. In particular, a high temperature, high velocity, 
sustained torch of hot gases is produced in a secondary, combustion 
chamber and introduced into the main combustion chamber through a torch 
tube. The primary characteristic of such torch ignition is that it assures 
starting and operating the engine under various adverse conditions, such 
as in very cold weather, with poor fuel/air mixture, and completely 
eliminates the use of sparkplugs, glowplugs, or the like for starting. 
Additionally, ignition delay is reduced to a minimum. The fuel can also be 
injected slowly (at a controlled rate) without causing or enhancing 
detonation within the main combustion chamber. With the elimination of 
detonation, the maximum cylinder pressure can be limited (by slow 
secondary fuel injection) to a desired value, thus permitting construction 
of a much lighter engine block and cylinder head. Lower operating 
temperatures also reduce production of oxides of nitrogen. 
The novel engine apparatus of this invention also involves a unique intake 
and exhaust system, the intake system producing a high degree of swirl 
within the main combustion chamber and for excellent mixing of the first, 
lean fuel and air mixture therein. The high swirl also readily assists in 
establishing a stationary flame front in the combustion chamber by 
continuously bringing new fuel/air through the combustion flame front 
thereby achieving substantially improved combustion. As a result of the 
foregoing, very few adverse combustion products such as soot are produced 
while providing a more efficient engine which can be operated at an 
optimum compression ratio and, therefore, is substantially quieter in 
operation. A more complete combustion also greatly reduces the emission of 
hydrocarbons, carbon monoxide, and other products of incomplete combustion 
to minuscule amounts. 
Since the engine of this invention is readily adaptable to either a 
two-stroke or four-stroke cycle, higher rotational speeds can be achieved 
since (a) the engine has lighter pistons (lighter pistons being possible 
because of a lack of a high peak pressure, crankshafts, etc.), and (b) the 
cushioning effect of the compression at the end of each stroke as the 
piston reaches TDC. 
The higher valving efficiency of this invention permits use of a two-cycle 
engine while concurrently permitting the power stroke of the engine to be 
of longer duration (approximately 135 degrees versus approximately 120 
degrees or less in the conventional engine) and also allows faster 
movement of gaseous products in and out of the engine, which thereby 
increases horsepower and efficiency of the engine. Additionally, this 
engine does not require an electronic ignition system which is frequently 
susceptible to failure and requires frequent maintenance, and as a result, 
the engine will be more reliable. The engine of this invention is also 
more reliable under engine of this invention is also more reliable under 
various adverse weather conditions, poor fuel/air mixtures, and the like. 
Referring now to the drawing, the novel torch ignition apparatus of this 
invention is shown generally at 10 and includes an engine block 12 with a 
piston 14 cooperating in a cylinder 16. Cylinder 16 is surmounted by a 
cylinder head 18 thereby forming a combustion chamber 20 between piston 14 
and cylinder head 18. 
A torch piston 22 is coaxially mounted to piston 14 by means of a threaded 
boss 23 and slidingly cooperates in a torch cylinder 26. A torch ignition 
chamber 30 is formed between the upper end of torch piston 22 and a 
spring-biased, torch cylinder head 24. Torch cylinder head 24 is a 
spring-biased piston in that it is movable in the upper end of torch 
cylinder 26. However, to avoid confusion, and also since it cooperates 
with torch ignition piston 22 to form torch ignition chamber 30, it will 
be referred to throughout as torch cylinder head 24. An annular flange 38 
around the upper end of torch cylinder head 24 is adapted to be placed in 
abutment with an upper end 27 of torch cylinder 26. A spring 34 serves to 
bias torch cylinder head 24 downwardly so that the flange 38 is brought 
into abutment with upper end 27 of torch cylinder 26. Spring 34 is 
maintained within a spring housing 32 mounted to upper end 27 of torch 
cylinder 26. A spring block 36 is threadedly engaged with spring housing 
32 and serves to adjustably maintain the desired compression on spring 34. 
An ignition torch exhaust port 42 is formed in a wall of torch cylinder 26 
in the vicinity of torch ignition chamber 30 and is in communication with 
a torch tube 40 extending downwardly into fluid communication with 
combustion chamber 20 through a torch ignition port 56. An intake port 50 
is formed in the wall of torch cylinder 26 and provides fluid 
communication for a torch fuel/air mixture 52 introduced through an intake 
tube 28 (shown broken for ease of illustration). A check valve is formed 
in intake tube 28 and consists of a valve body 44 cooperating between a 
valve intake 48 and a valve exhaust 46. 
Cylinder head 18 is mounted to the upper end of engine block 12 by means of 
a plurality of cylinder head pillars 66 formed coextensively with the wall 
of cylinder 16 and fabricated as a plurality of spaced pillars around the 
upper periphery of cylinder 16. Cylinder head 18 includes a cylindrical 
wall 68 over which a hoop valve 60 slidingly cooperates. Wall 68 extends 
upwardly a predetermined distance to provide a sealing relationship when 
hoop valve 60 is raised upwardly. Intake ports 70 are formed as the spaces 
between cylinder head pillars 66 and are formed as a plurality of intake 
ports around the periphery of combustion chamber 20. Intake ports 70 are 
closed by downward movement of hoop valve 60 with the space represented by 
intake ports 70 being occupied by valve inserts 64 mounted to the internal 
face of hoop valve 60. The openings of valve intake ports 70 are covered 
by valve inserts 64 which are fabricated from a suitable high temperature 
material such as zirconium oxide, silicon nitride, or other conventional, 
ceramic, or cermet materials having the desirable high temperature 
properties. 
Cut-outs (not shown) in the internal face of hoop valve 60 represent the 
spaces between valve inserts 64 and provide the necessary opening in hoop 
valve 60 for cylinder head pillars 66 when hoop valve 60 is lowered 
downwardly thereover. Hoop valve 60 includes a plurality of valve rings 62 
to accommodate sealing relationship between hoop valve 60 and wall 68. 
Hoop valve 60 further includes an outwardly extending flange 61 which is 
adapted to sealingly engage a plurality of compression rings 76 formed in 
the upper periphery of cylinder 16. 
Advantageously, hoop valve 60 provides a substantially increased intake 
area for intake ports 70 even though a substantial portion (up to 1/2) of 
the potential surface area is blocked by cylinder head pillars 66. 
Decreasing the width of cylinder head pillars 66 relative to the width of 
intake ports 70 would, selectively, provide a corresponding increase in 
the surface area of intake ports 70. A simple mathematical calculation 
will demonstrate that there is a several-fold increase in intake port area 
provided by a hoop valve 60 versus the conventional poppet valve systems. 
The upper periphery of engine block 12 is formed as an annular rim 75 
having a plurality of guide vanes 74 mounted thereon. Rim 75 also serves 
as the basal support for an intake housing 92, intake housing 92 enclosing 
all of the upper portion of the ignition apparatus 10 of this invention 
and also serving as a covering to provide an intake plenum 90 as will be 
set forth more fully hereinafter. Intake housing 92 also includes an 
intake port (not shown) to provide fuel/air from the intake manifold (not 
shown) into the intake plenum 90. 
The lower portion of cylinder 16 includes a plurality of exhaust ports 80. 
Exhaust ports 80 communicate with an exhaust plenum 82 formed as an 
annular chamber surrounding the lower portion of cylinder 16. Exhaust 
ports 80 are fabricated as a plurality of openings between columns 78, 
columns 78 serving to provide the necessary interconnection for cylinder 
16 and also including coolant channels 17 therethrough. The arrangement of 
exhaust ports 80 is also novel in that a greatly enlarged exhaust port 
area is exposed upon the downward movement of piston 14 thereby permitting 
the rapid clearing of exhaust products from combustion chamber 20 as will 
be set forth more fully hereinafter. The only limiting factor therein is 
determined by the width of pillars 78 and the height of exhaust ports 80. 
Cooling for engine block 12 is provided through coolant channels 17 
surrounding cylinder 16 and coolant channel 19 in cylinder head 18. 
In operation, the engine apparatus 10 of this invention is configurated for 
two-cycle operation with a flowthrough of gaseous products from the upper, 
intake ports 70 through the combustion chamber 20 and out the lower, 
exhaust ports 80 into the exhaust plenum 82. However, torch ignition will 
work equally well on four-cycle engines. Intake fuel/air mixture is 
introduced into intake plenum 90 from the intake manifold (not shown) and 
directed by vanes 74 through intake ports 70. The orientation of vanes 74 
is such that the fuel/air mixture in combustion chamber 20 has imparted 
thereto a counterclockwise swirl. After the fuel/air mixture has been 
introduced into combustion chamber 20, hoop valve 60 is closed by being 
moved downwardly into sealing relationship with valve rings 76. The upward 
movement of piston 14 then compresses the fuel/mixture in combustion 
chamber 20. 
A second fuel/air mixture 52 is introduced through port 50 into torch 
ignition chamber 30. The second fuel/air mixture 52 is of the desired 
optimum fuel and air mixture to support combustion upon compression as in 
a compression ignition system. Upward movement of piston 14 
correspondingly moves torch piston 22 upwardly to compress the second 
fuel/air mixture 52 in torch ignition chamber 30 so that ignition of the 
same is obtained. The desired ignition point for the second fuel/air 
mixture in torch ignition chamber 30 is selectively obtained by adjustably 
setting the tension on spring 34 through upward or downward movement of 
spring block 36. The ignition of the second fuel/air mixture 52 in torch 
ignition chamber 30 results in an explosion and a corresponding raising of 
torch cylinder head 24, as illustrated. The hot combustion products are 
allowed to escape through port 42 into torch tube 40 and out torch outlet 
56 into combustion chamber 20. 
During the downstroke, torch piston 22 again drops below port 50 to permit 
the pressurized, second fuel/air mixture 52 to be introduced into torch 
ignition chamber 30. Correspondingly, torch cylinder head 24 and, more 
particularly, rim 38 is forced downwardly against upper end 27 by spring 
34 so that torch cylinder head 24 closes outlet port 42. Importantly, the 
tension on spring 34 is adjustably predetermined so that a preselected 
pressure is obtained on the second fuel/air mixture in torch ignition 
chamber 30 to cause a selectively timed ignition of the same under 
pressure. The resultant increased pressure following ignition again raises 
torch cylinder head 24 against spring 30 exposing exit port 42 to thereby 
permit the hot combustion products to be vented through torch tube 40 and 
out through torch port 56. 
Supplemental fuel is simultaneously supplied through fuel conduit 54 in 
conjunction with the hot combustion products through torch port 56. Torch 
port 56 is mounted adjacent fuel conduit 54 and tangentially to combustion 
chamber 20 thereby enhancing swirling action inside combustion chamber 20. 
The swirling action imparted to the first fuel/air mixture by vanes 74 
brings a continuous supply of compressed, first fuel/air mixture into 
contact with hot combustion products and supplemental fuel thereby 
providing ignition of the combined fuel/air mixture within combustion 
chamber 20. Importantly, the first fuel/air mixture in intake plenum 90 
has an inadequate fuel ratio therein to support combustion in the absence 
of supplemental fuel. The supplemental fuel introduced through conduit is 
heated above its combustion temperature so that contact with oxygen in the 
first fuel/air mixture results in a combustion not only of the 
supplemental fuel but also the initial fuel in the first fuel/air mixture. 
Advantageously, the introduction of hot combustion gasses through port 56 
provides a sustained ignition of the first and second fuel/air mixtures 
for a more thorough combustion of the same during the time span that 
piston 14 is moving upwardly toward top dead center and during the initial 
stages of the power stroke. This is a distinct advantage over the 
relatively short ignition span of a compression ignition engine and even 
of a conventional spark-initiated combustion cycle and also permits 
control of maximum combustion chamber pressure, keeping it lower than in 
current engines, thus permitting lighter engine construction. 
Additionally, since diesel-type fuels tend to burn extremely fast or even 
detonate when carbureted with air (oxygen), the first fuel/air mixture 
must have quite a low percentage of fuel. However, by adding about 10 
percent water, either to the fuel or mixed with the induction air, the 
percentage of fuel in the first fuel/air mixture can be increased without 
incurring detonation for better burning and, subsequently, more complete 
combustion resulting in a higher efficiency. 
Thereafter, piston 14 continues its downward stroke until the exhaust ports 
80 are exposed with a resultant expansion of the exhaust products 
therethrough into exhaust plenum 80. Simultaneously, hoop valve 60 is 
raised upwardly permitting the pressurized first fuel/air or water/air 
mixture in intake plenum 90 to be swirled by vanes 74 into combustion 
chamber 20. The inrush of first fuel/air mixture into combustion chamber 
20 also serves to sweep all of the residual exhaust products through 
exhaust ports 80. Reversal of the direction of piston 14 to the upward 
direction again covers exhaust ports 80 and the timed closure of hoop 
valve 60 allows piston 14 to compress the first fuel/air mixture in 
combustion chamber 20, thereby repeating the ignition cycle. 
The invention may be embodied in other specific forms without departing 
from its spirit or essential characteristics. The described embodiments 
are to be considered in all respects only as illustrative and not 
restrictive and the scope of the invention is, therefore, indicated by the 
appended claims rather than by the foregoing description. All changes 
which come within the meaning and range of equivalency of the claims are 
to be embraced within their scope.