FLOW COUPLED ARC DISCHARGE IGNITION IN AN IC ENGINE

Abstract: An improved ignition-combustion system for internal combustion engines comprising a compact combustion chamber zone (4) mainly in the engine cylinder head (6) mainly under the exhaust valve (8) and with large air-squish zones (124a, 124b) formed at the edge of the combustion zone which produce colliding squish flows (2, 2a, 3a, 3b) with high turbulence (3c) at the center of the combustion zone, with one or two spark plugs (12a/118, 12b/18a) located at the edge of the combustion zone within the high squish zones, resulting in a combined ignition and combustion system of colliding-flow-coupled-spark discharge (CFCSD), with the ignition employing high energy flow-resistant ignition sparks which move under the influence of the squish flow towards the central turbulence region as the piston nears engine top center, to produce rapid and complete burning of lean and high EGR mixtures for best engine efficiency and lowest emissions.

Description: FLOW COUPLED ARC DISCHARGE IGNITION IN AN IC ENGINE FIELD OF THE INVENTION This invention relates to spark ignition internal combustion (IC) engines and to the improved ignition and combustion of air-fuel mixtures in IC engines through the faster and more vigorous combustion of the air-fuel mixture brought about by high air-flows of both the directed bulk flow type and of the more random turbulent flow type produced m a highly compact combustion zone of the combustion chamber In particular, the invention relates to combustion systems which can simultaneously produce high directed bulk flow at the spark plug site at the time of igmtion and more random and intense turbulent flow inside a compact combustion chamber dunng the time of ignition and combustion to speed-up the burn In particular the invention relates to a system which involves the design and control of high flows, high energy ignition spark discharges, and resulting intense flame kernels, all three of which interact among themselves in ways that produce a very rapid and controlled burn in a comp act combustion zone for low NOx emissions and high engine efficiency through lean and high EGR (exhaust gas recirculation) combustion at a high engine expansion ratio BACKGROUND OF THE INVENTION AND PRIOR ART High flows are used in IC engines to improve engine efficiency and emissions through rapid combustion of lean and high EGR air-fuel mixtures These flows are either swirl tumble (vertical vortex), or squish The predominant type of flow used is swirl, especially for the purpose of generating mixture stratification at the spark plug site to allow for ignition of lean mixtures Numerous examples of these exist, especially with Japanese manufacturers such as Toyota, Mazda, and others To a lesser extent tumble flow is used, as in the Mitsubishi Vertical \ ortex (MW) engine now in production Squish is rarely used for efficiency and emissions improvement It is used principally in racing and performance for speeding up the combustion of high speed engines Moreover, such squish flow is not used in conjunction with the spark except in my U S patent No 5,517,961, where it is used in conjunction with a high energλ flow- resistant spark to help spread the spark and speed up the burn SUMMARY OF THE INVENTION In this patent application is disclosed an ignition and combustion system based on the interaction of two generic types of flows which can be generated in an engine cylinder. One is bulk flow which can be produced at the spark plug site at the time of ignition to interact /couple to the spark discharge to direct and spread the spark, along arc runners if practical; the other is turbulent flow which can be produced by impinging squish flows directed towards each other as colliding flows reinforced by the piston compression of swirl or tumble flow, into which the spark discharges and resulting initial flame kernels can move in the form of flow-coupled-spark- dischargesofthetype disclosed in myU.S. patentNo. 5,517,961. Preferably the more random turbulent flows (colliding flows or piston converted swirl or tumble) occur in ahighly confined region or zone of the engine combustion chamber defined by the piston or other movable element nearing top center (TC) and the walls of the combustion volume, made up, for example, of the engine cylinder wall and cylinder head in a conventional piston engine. Preferably, the combustion zone is located partly or mostly under the exhaust valve, or in some other high temperature zone of the combustion chamber, with the bulk flow occurring at one or more edges of the turbulent zone to move and direct one or more high energy flow-resistant spark discharges into the mixture turbulence to produce a rapid and complete combustion of an air-fuel rnixture without engine knock or other deleterious effect. In this way, very lean and high EGR mixtures can be burnt rapidly to produce high engine efficiencies and low emissions, especially low NOx emissions whose formation depends on both temperature and time, both of which are reduced. With the use of one spark plug and two valves, the spark plug is preferably located between the valves disposed preferably along or near a center diameter line of the engine cylinder if bulk flow is present at the spark plug site, e. g. if the combustion chamber is under the exhaust valve; or the spark plug may be placed offset of the valve center line at a high bulk flow region if the combustion chamber includes part or most of the region under both valves, either in the cylinder head or in the piston as has been disclosed in my U. S . Patent No.5,517,961. If two spark plugs are used, with one or two intake and one exhaust valve, more options are available for achieving spark coupling to the bulk flow. The spark plugs can be disposed on either side of the exhaust valve under which is defined the main combustion chamber and region of turbulent flow, with two inward moving bulk flows (BFs) occurring at the two spark plug sites for two flow- coupled-spark discharges (FCSD), producing a more elongated combustion zone and even more rapid burning of the air-fuel mixture. Placing the combustion chamber under the exhaust valve improves the engine' s lean burn capability and reduces its knock susceptibility, especially if the end-gas region of the combustion chamber is narrow and subj ect to cooling. These are preferred embodiments of the invention. Furthermore, in these preferred embodiments, the piston is flat, or slightly cupped under the combustion zone, with thermal barrier coatings used to limit heat transfer to the piston. Near top center (TC), i.e. within 30° of TC, the bulk flow (BF) in the spark gap is produced by the piston which induces squish flows which move towards each other and collide to, in turn, produce an intense turbulence in the combustion zone (a colliding turbulence generating flow, or CTF). Advanced of TC, the bulk flows in the spark gap are produced by swirl or tumble, or by a combination of these which are converted by the piston motion near TC to turbulent flows to produce a very rapid and vigorous burn. When squish flow dominates this process, it can be viewed as a colliding-flow-coupled-spark-discharge (CFCSD), versus simply an FCSD for swirl or tumble. Both are made up of high energy, flow-resistant, flow-coupled- spark-discharges which move into, and/or produce initial flame kemels which move into the high turbulence region (of small eddies),produced by colliding flows or piston compression converted bulk flow which help homogenize the mixture and help spread the flame. In another preferred embodiment, the inlet air flow is used to reinforce the combustion process. Such an inlet flow can be a vertical vortex flow which reinforces one or more of the squish induced flows, especially at advanced ignition timings when squish is low. Alternatively, the combustion chamber and inlet valve, one or two inlet valves, and intake runners may be designed to produce swirl flow to reinforce the squish flows. For the ignition, preferably triangular distribution spark with a peak amplitude above 200 milliamps (ma) is used, say 300 to 800 ma, operating in part in the flow-resistant arc discharge mode (versus glow discharge) which can sustain flow velocities of 20 meters/sec (m/s) or greater without spark segmentation, as disclosed in my U. S. Patent No.5,517,961 , to produce a very fast ignition and combustion of the mixture. In another preferred embodiment, the spark may be fired to the piston given the ability to operate the engine with ignition near top center, between 5 ° and 30° before top center (BTC), with special electrode tips acting as arc runners which improve coupling of the spark to the flowing mixture. Special design halo-disc spark plugs of my U.S. Patent No.5,577,471 canbe used, or similar plugs with large thin discs at their ends to define a runner along which the spark can move under the influence of the flow. Preferably, erosion resistant material such as tungsten- nickel-iron is used for the spark plug electrode material. In another preferred embodiment, the valve timing is selected to make best use of CFCSD. This includes setting the intake valve opening preferably near TC (a few degrees before TC) for the cases with no combustion volume under the intake valve. The exhaust valve closing is preferably set at or before TC (BTC), e.g. 0° to 30° BTC to trap the last part of the exhaust for both high internal exhaust gas residual fraction to eliminate or ininimize the need for EGR as well as minimize hydrocarbon (HC) emissions since the last part of the burnt mixture to be exhausted contains proportionally more unbumtHC. Havingthe combustion chamber under the exhaust valve can help the process of trapping the (last) part of the exhaust corning from the ring clearances and other chamber crevices which have the most unburnt HC emissions, to be burnt in the subsequent ignition and combustion firing cycle. In another preferred embodiment using preferably a high energy, high efficiency, flow resistant ignition spark, an improved combustion chamber of the squish type is used which has the combustion chamber in the cylinder head under both the intake and exhaust valves which are centrally located and made as small as practical to allow for large squish flows, especially in a direction transverse to the line joining the two valves, with one or two spark plugs located in the squish zone. Preferably, the intake valve is recessed so that in its fully open condition it clears the top of the piston when positioned at top center (TC), and likewise for the exhaust valve which is preferably even further recessed. This represents a free-wheeling system which will not damage the engine should the timing belt break (in the preferred overhead cam configuration). In such a design several advantages result, including thorough mixing of the air-fuel mixture due to the squish induced colliding flows, rapid burning of the air-fuel mixture, especially of high dilution, high exhaust gas recirculation (EGR) and high air-fuel ratio (AFR) mixtures which produce high efficiency and low emissions, and high internally retained exhaust gas, especially that located at the further regions of the combustion chamber, e. g. in the ring crevices where most of the unburnt hydrocarbons are formed which can be re-burned in the subsequent ignition firing cycle. Preferably, a squish land exists around the entire periphery of the piston so that the end gas can be cooled and therefore allow for higher compression ratio operation. More ideally, the disclosed ignition/combustion/flow systems are used in aMiller Cycle engine or variable- valve-timing engine with high expansion ratios of 10 to 12 to one or higher and delayed intake valve closing (for effective lower compression ratio). If practical, the systems disclosed are used in the virtual 3 -stroke engine disclosed in my U. S . patent SN 5 ,454,352 where inFIG.7ofmatpatentisshownaprefertedembodimentofιhevirtxιd3-sfrokeengmemtheform of an opposed piston with two valves and two spark plugs. Preferably, the valves are smaller which is more practical with the use of a turbocharger to create a more compact combustion chamber in the cylinder head and higher squish at the spark plug sites. The low compressive friction makes this engine easy to start and ideal for the forthcoming dual rail 14/42 volt battery systems with integrated starter/generator (ISG). It also makes it very efficient and of very low emissions. NOx emissions are low due to the low adiabatic heating of the intake air on compression and low overall temperatures due to the long expansion stroke (preferably of 12 to one or higher expansion ratio). Because ofme inherently high efficiency (BSFC, or brake specific fuel consumption), especially at part load, combustion can be somewhat delayed, as in the diesel engine, to limit peak combustion temperatures below the level where significant NOx emissions are formed. This engine can be operated either premixed, i.e. indirect inj ection or carbureted, or direct in-cylinder injection, either moderate pressure injection as in the GDI engine, or higher pressure injection as in the diesel engine. Other improvements of the ignition-engine system include improved forms of dual ignition, especially of the low inductance type, high energy, high efficiency inductive ignition, which use capacitance type plugs, if practical, with erosion resistance electrode materials. Also, such low inductance type coils can be improved by having magnetically uncoupled secondary high voltage windings to aid in reducing an over-voltage coil end effect associated with low resistance high efficiency coil windings. Other means for producing CFCSD are possible, including in other types of engines, such as two-stroke, rotary, gas direct inj ection, and others, the main inventive principle disclosed herein being the combination and interaction of flow-coupled-spark-discharge, especially ofhigh-energy high-efficiency flow-resistant sparks with turbulence generating flows, especially in a compact, hot region of the combustion chamber of a spark ignition engine. BRIEF DESCRIPTION OF THE DRAWINGS FIGS, la and lb are drawings of air-flow vectors and flow interactions in a compact engine combustion chamber (combustion zone) that characterize a key feature of the invention. FIGS. 2a and 2b are approximately to-scale drawings of the side and top view of a combustion chamber with a preferred embodiment of the invention with central combustion zone under a central exhaust valve, and two intake valves and spark plugs located in the squish zones . FIG. 3 is a top view similar to that of FIG. 2b but with one intake valve. FIGS.4ato 4c show various arc runners for the spark to move along under the action of bulk flow, such as squish flow. FIG.5 is an approximately to-scale side view drawing through a line intersecting the two spark plugs of FIGS. 2b and 3 with arc runner construction. FIGS. 6a and 6b are two types of spark plug firing ends for ignition firing to the piston. FIG. 7 is a more idealized embodiment of FIG. 5 with special spark plugs located on an electrically insulating cylinder head, achievable in a two stroke engine or other engine where valves are absent from the cylinder head or highly confined in the head. FIG. 8 is apreferred embodiment of a direct injection version ofthe invention with the fuel injector located at the edge of a combustion zone under the exhaust valve across from a single spark plug in the squish zone. FIG. 8a is a possible side view ofthe embodiment of FIG. 8 during the intake process. FIG. 8b is a view of FIG. 8 without the fuel injector, with the piston near top center. FIGS.9a and 9b is atop and side view drawing of a combustion chamber with two intake valves and one exhaust valve with the combustion zone under the exhaust valve and a single spark plug located at a more central point at the edge ofthe combustion zone. FIGS. lOaand 10b is a more conventional pent-roof chamber with combustion zone both in the head and piston under the single intake and exhaust valves and two spark plugs at the edge ofthe combustion zone in the squish zones, FIGS . l l, lla,l lb,l lc and 11 d are partial side views and a top view of a combustion chamber with one small intake and exhaust valve centrally located and with the compact combustion zone under the two valves, and with two spark plugs located at two opposite sides of the combustion zone in regions ofhigh squish flow, and a central fuel inj ector shown as an option in FIGS, l ie and lid. FIGS. 12a, 12b, 12c, and 12d are approximately to-scale partial top and side views of a combustion chamber ofFIGS. 11, 11a and 11 b detailing and emphasizing certain aspects ofthe combustion chamber for improved operation. FIG. 13 is a cam timing diagram showing a preferred valve timing for the present invention. FIG. 14 is a design ofthe ignition coil output structure for providing dual ignition firing to the piston ofthe present invention in which the return path for the electrical current is entirely though the spark plugs and not through the engine block. FIG. 15a is a partial, not to scale, top view layout of two side-by-side engine cylinders in a multi-cylinder engine, with FIG. 15b depicting a preferred approximately to-scale combustion chamber with dual ignition ofthe present invention in which the valves are lined up approximately in the length axis ofthe engine. FIG. 16a is a partial, not to scale, top view layout of two side-by-side engine cylinders in a multi-cylinder engine, with FIG. 16b depicting a preferred approximately to-scale combustion chamber with dual ignition ofthe present invention in which the valves are lined up transverse to the length axis ofthe engine. FIGS. 17a, 17b and 17c are partial top views oflow inductance, high energy ignition coil structures with open magnetic paths suitable for application where greater than 12 volts power is available and designed for having two high voltage towers suitable for dual ignition. FIG 18 is atop partial view of an open E-type low inductance, high energy igmtion coil with open magnetic path with some lower turns loosely coupled secondary windings constituting a series of secondary circuit external leakage inductors, usable also in the designs ofFIGS. 17a and 17b, for reducing the end effect voltage overshoot. FIG. 18a is an equivalent circuit ofthe coil of FIG. 18 used in an inductive ignition circuit. DESCRIPTION OF PREFERRED EMBODIMENTS FIGS, la and lb are drawings depicting, in schematic form, flow vectors and flow interactions that lead to the principles of this invention. Region 1 denotes the region ofthe ignition source which can be the gap between spark plug electrodes and the region adjacent to the electrodes into which the spark discharge moves by the action of the air-flow indicated by flow vectors 2, representing bulk flow, BF, to produce what has been described as a flow-coupled- spark-discharge, FCSD. Region 3, shown with small partially circular curves with arrows 3c, indicating tiny turbulent eddies, is the region of high turbulence generated by the approximately equal and oppositely directed colliding flows 3 a and 3b, producing what has been designated as CT. The spark discharge region 1 preferably penetrates the turbulence region 3 at approximately right angles to the direction of the colliding flow vectors 3a and 3b, as shown. A second ignition source region can be equally well placed at the right side ofthe turbulence region 3, as is indicated with reference to FIGS. 2a, 2b, and other figures. FIGS. 2a and 2b depict in partially schematic, approximately to-scale drawings, side and top views of a combustion zone 4 of a conventional piston engine contained within a larger combustion chamber volume defined by a piston 5, cylinder head 6, and cylinder sleeve 7. The term "combustion zone" 4 shall be used to indicate the region where the major and first part ofthe combustion, or burning, occurs as distinguished from "combustion chamber" which indicates the entire volume as already defined. In this preferred embodiment, the combustion zone 4 is shown located in a central part of the combustion chamber under an approximately centrally located exhaust valve 8 with exhaust tube 8a, with two intake valves 9a, 9b shown at one side of the combustion chamber with intake tubes 9c and 9d (allowing for the more central exhaust valve 8 and combustion zone 4 under the exhaust valve 8). In this preferred embodiment is shown two high voltage spark plug electrodes 10a and 10b disposed symmetrically at opposite ends of the elongated combustion zone 4 with ground electrodes 11a and l ib defined by small rises on the piston for producing spark discharges between the tips 10c and lOd of electrodes 10a and 10b and the piston surface regions 11a and l ib, i.e. ignition firing to the piston versus to conventional "J" ground electrodes. Spark firing tips 10c, lOd, 1 la, and 1 lb are made of erosion resistant material. For ease of visualization in FIG. 2a are shown, around the gap ofthe left-most electrodes lOc/lla, bulk flow vectors 2, and around the right-most electrodes lOd/l lb spark discharges 1 being directed by the bulk flow vectors into the turbulence region 3, understanding that both the bulk flow (2 and 2a of FIG. 2b) and spark discharges (1 and la of FIG. 2b) are present around both spark gaps. In this preferred embodiment, a central combustion zone is provided with high bulk flow at the plug sites and high turbulence inside the combustion zone. When used with a high energy ignition system with flow resistant sparks (direct current arc discharge mode) of long duration, i.e. of order of magnitude one millisecond, one can achieve a very rapid bum of a dilute air-fuel mixture, diluted with excess air (lean mixture) andor with retained exhaust gas (high exhaust residual or EGR), for very high efficiency and low emissions. Preferably, a high and optimum compression ratio (CR) is used, e.g. between 10:1 CR and 15:1 CR, for best efficiency, which depends on several factors including fuel type used and heat transfer to the combustion chamber surfaces which increases with compression ratio (and can be reduced with thermal barrier coatings). More ideally, a Miller type cycle is used with late intake valve closing of 60° after bottom center (ABC) or greater, and a high expansion ratio (ER) of 12 to one or higher. FIG. 3 is an approximately to-scale, partially schematic top view of a spark ignition piston engine representing a preferred embodiment of CFCSD ofthe type ofFIGS. 2a and 2b except that it has one intake valve 9 instead of two. Like numerals represent like parts with respect to the earlier figures. In this preferred embodiment, the spark plugs 12a and 12b are shown (instead of the electrodes), understanding that the spark plugs can be located at an angle to the vertical as is depicted in FIG. 5, versus vertical as shown in FIGS. 2a, 2b. This design is especially useful for use with a centrally located single overhead cam 13, as depicted, with the two spark plugs located with an angle to the vertical to help clear the cam as well as to better position their tips in the combustion chamber. Another embodiment is to have two smaller intake valves, one at the present location and the other at the opposite side, 180 degrees away. For better coupling ofthe spark discharge to the flow, an electrode geometry that acts in part as an arc runner may be practical. FIGS. 4a to 4c show a pair of electrodes 10 and 11 with firing tips 13 and 14, the electrodes subtending progressively smaller included angles from FIG. 4a to 4c. Flow vectors 2 are shown emanating from the spark gaps 15 as are spark discharges 1 which move along the runners 10/11. It can be appreciated that the smaller the included angle between the electrodes, say from 30° to 90°, the more directed is the flow and the less sfretelύng there is ofthe spark discharge as it moves along the runners 10/11 per unit motion, making FIG. 4c the more preferred embodiment for electrode geometry which can produce greater motion of e spark discharges 1 and hence greater penetration of the spark into the turbulence region 3. FIG. 5 shows an embodiment of the spark plug electrodes of FIGS. 2b and 3 to produce better spark penetration and size. FIG. 5 is a partially schematic, approximately to-scale side view drawing through a line intersecting the two spark plugs of FIGS. 2b and 3, representing a preferred embodiment of CFCSD. Like numerals represent like parts with respect to the earlier figures. In this embodiment are shown spark plugs 12a and 12b, preferably 18 mm, with large diameter, thin, erosion resistant electrodes 10c, lOd disposed at right angles to the plug axis forrning a small included angle (approximately 45°) with the piston top (shown close to TC) to produce a preferred arc runner of FIG. 4c (versus FIG. 2a which represents arc runners of FIG. 4b). As in FIG. 2a, the flow vectors 2 are shown on the left and the spark discharges 1 on the right for ease of visualization, both moving into the turbulence region 3. The spark gaps 15 are made with small rises 1 la and 1 lb on the piston surface, which represents the second arc runner combination 11 of FIGS. 4a to 4c. Systems involving firing to the piston have been disclosed in several of my prior patents, including U.S. patents Nos. 4,774,914 and 4,841,925 among others. Also, many electrode types have been disclosed, two of several possible types being depicted in FIGS. 6a and 6b. FIGS. 6a and 6b are approximately to-scale, partial side view drawings of two spark plug ends firing to the piston 5. Like numerals represent like parts with respect to the earlier figures. In FIG. 6a the center conductor 16 is preferably copper extending axially into the combustion chamber near the piston surface at the TC point with a coating 10 (representing the arc runner) of erosion resistant material, such as tungsten-nickel-iron. The runner geometry conforms to that of FIG. 4b (90° included angle). In FIG. 6b the arc runner is erosion resistant disc 17 at right angles to the plug axis shown located slightly away from the end surface of insulator 18 (versus on the surface of the insulator, as in FIG. 5). These piston firing embodiments produce better spark penetration towards the center ofthe combustion zone. FIG. 7 is a more idealized embodiment of FIG. 5 which can be attained in a two-stroke engine where valves are absent from the cylinder head or rotary type valves are used (with a four stroke engine also). The engine cylinder is a partially schematic, approximately to-scale side-view drawing representing another preferred embodiment of CFCSD. Like numerals represent like parts with respect to the earlier figures. In this preferred embodiment, the cylinder head is shown made of insulating thermal barrier material (as in an adiabatic engine), with a central triangular cross- section combustion zone 4, on the two angled surfaces of which are placed large, flat, thin electrodes 17a, 17b connected to high voltage terminal 16a and 16b respectively. In this embodiment, extremely effective and large arc runners between 17a and 17b and the piston surface are achievable for extensive penetration of the arc discharge into the turbulence region 3 of the combustion zone 4. However, to accomplish this a high energy ignition with long duration flow- resistant sparks ofthe types I have disclosed elsewhere is preferred. Also, special treatment ofthe piston surface is required, e.g. such as coating with erosion resistant material which can also be a relatively poor thermal conductor. In many applications arc runners are not practical, so more conventional spark gaps with proper oriented electrodes may be used (that do not interfere with the flow), or circularly symmetric spark gaps may be used. In such a case, the high energy-flow coupled-spark-discharge (FCSD) acts as a stationary extended spark (a plume) through which the mixture flows under the action of bulk flow to produce flame kernels which move and grow as they enter the turbulence region 3 of the combustion zone 4. FIG.8 depicts such an embodiment. In FIG. 8 is shown a partially schematic, approximately to-scale top view of a piston IC engine representing a preferred embodiment of CFCSD with a single spark plug 12 and single intake valve 9, with the valves and spark plug placed collinear in an approximately central cylinder 7 diameter line with the spark plug in the center between the valves.Like numerals represent like parts with respect to the earlier figures. In this embodiment, the combustion zone 4 is located mainly under the exhaust valve with small piston clearance for the intake valve to allow for squish bulk flow 2 at the spark plug site. The use of a rotary intake valve would obviate any concern associated with the preferred small clearances between the valve and piston. Shown also in the figure is a fuel injector nozzle 18 preferentially located across from the spark plug, representing a gas-direct-injection (GDI) version ofthe present invention allowing for the use of various fuels. The injector injects the fuel spray 19 into the turbulence region where it becomes rapidly mixed and homogenized for clean burning and prevention of spark plug fouling (a problem with spark ignition diesels). A partially schematic, side view, approximately to-scale drawing of a cylinder of the engine of FIG. 8 with the piston near the end ofthe intake stroke is shown in FIG. 8a. The intake air 20 is shown entering in a vertical circular form representative of tumble flow which can reinforce the flow 2 at the spark gap 21 of FIG. 8b, which represents the engine of FIG. 8 a with the piston at the end ofthe compression stroke near TC. Like numerals for both FIGS. 8a and 8b represent like parts with respect to the earlier figures. In FIG. 8a is shown the alternative possible fuel nozzle 18 (not shown in FIG. 8b representing the more conventional engine). FIG. 9a shows a partially schematic, approximately to-scale top view of a piston IC engine representing a preferred embodiment of CFCSD with a single spark plug 12 and two intake valves 9a and 9b, with the exhaust valve 8 and spark plug 12 placed collinear in an approximately central cylinder 7 diameter line. Like numerals represent like parts with respect to tie earlier figures. The shaded portion shown indicates the sloping region ofthe combustion zone from its peak just under the exhaust valve 8. FIG. 9b is a partially schematic side-view drawing ofthe engine cylinder of FIG. 9a, with like numerals representing like parts with respect to the earher figures. As in most ofthe drawings, the piston is shown near TC in the compression ignition firing stage. In this embodiment, the combustion chamber zone 4 is under the exhaust valve 8 and the spark gap is in a high squish zone as required. FIG. 10a and 10b indicates the combustion chamber of a more conventional type pent-roof engine with the combustion zone shared between the cylinder head 6 and the piston 5, but mostly in the piston. Two spark plugs 12a and 12b are shown as one of several options (one spark plug being the more conventional option). The spark plug tips may be more conventional or similar to those of FIGS. 5 and 6b which can act as arc runners as already discussed. Like numerals represent like parts with respect to the earlier figures. An interesting and relevant variation of the combustion chamber of FIGS. 10a and 10b is that of Jim Feuling most recently reported in the February 1999 issue of Hot Rod magazine in an article entitled "Lean, Mean, and Clean, Part One". The combustion chamber is a Figure 8 combustion chamber in d e cylinder head under the intake and exhaust valves with the spark plug in the middle. This design creates high turbulence in the combustion chamber except that the volume is larger (combustion zone is under both valves). Moreover, since the spark plug is in the middle and not at the edges ofthe combustion zone, as in FIGS. 10a and 10b, and does not use a high energy, flow resistant spark, it is not subject to bulk flow, as is disclosed in the present invention, to spread the spark discharge and create a rapid initial burn. In the preferred embodiments of this invention, the ignition and combustion stage is so fast that ignition can occur near TC, even for dilute mixtures (which permits piston firing as one alternative for generating and locating the spark, although it is not required). Preferably, ignition occurs as close to 10° BTC as practical since the inward directed squish flow (bulk flow) is at its maximum near that ignition timing. Under such conditions, the very powerful and rapid ignition and bum means that highly dilute mixtures can be burnt effectively for low exhaust emissions and high fuel economy. Preferably mixture dilution is produced internally (versus EGR) by valve timing, i.e. by advancing the exhaust valve closure, although it may be supplemented with EGR and with excess air (lean mixtures). The very fast bum of dilute mixtures results in very low NOx emissions due to both the faster bum (less time for NOx to form) and lower peak combustion temperatures, while the higher trapped exhaust results in lower HC emissions. Also the bulk flow and colliding flow turbulence generation produces very good air-fuel mixing for a more homogen- eous mixture for lower HC emissions and higher engine efficiency. Rapid bum also allows for higher compression ratio and higher engine efficiency. FIGS . 11, 11a, l ib, lie and lid depict various partial views of an engine combustion chamber in which the compact combustion zone 110 is contained in the cylinder head under both the intake and exhaust valve of preferably small size and with two spark plugs in the combustion chamber located in the region of high flow. FIG. 11 depicts a top part of a spark plug 118 with a preferred capacitive spark plug boot 117, which may also be built into the spark plug to provide atypical capacitance of 30 to 60 picofarads. The combustion chamber zone 110 disclosed has preferably intake valve 121a and exhaust valve 121b disposed parallel to the flat top ofthe piston 119b and are made as small as practical to provide squish zones 122a and 122b at the far edges as shown. The valves are located in line as close as is practical to each other, as shown in 1he top view ofFIG. 11a, where like numerals represent like parts with respect to FIG. 11. The combustion chamber zone 110 is in the head in an elongated region defined by a volume under each valve and open in the central region 123 between the two valves. Preferably, the larger volume is under the exhaust valve 121b, which is more recessed than the intake valve 121a, which is preferably recessed to just to miss the piston at its TC position when the intake valve is fully open (which can occur if the timing belt breaks for an overhead cam engine design). With further reference to FIG. 11 a, the design shown provides large squish areas 124a and 124b, as is also shown in FIGS, lib, lie, lid which are side-views disposed 90° from that of FIG ll. JjiFIGS. llb,llc, lid likenumerals represent like parts with respect to FIG.11. In both FIGS. 1 la , 1 lb,l lc, 1 Id are shown the possibility for inclusion of two, instead of just one, symmetrically placed spark plugs 118 and 118a, making preferably about 30° with the vertical and placed atthe inner edge ofthe large squish zones 124aand 124b so that the spark gaps 125a and 125b are subjected to high squish flow near TC during ignition, hi a preferred design, the squishlands 124aand 124b have a transverse length "1" approximately equal to orgreaterthan the transverse radial dimension "r" ofthe compact combustion chamber, as is shown to produce high squish at the spark plug site and high colliding flows in the compact combustion chamber zonel 10. This helps vaporize, by the radial intense flow prior to TC, any fuel that may have collected on the piston lands, and produces rapid rnixing ofthe air-fuel mixture in the combustion chamber zone 110 due to the turbulent colliding flows, even at low speeds where combustion is normally poor. The piston to head clearance in the land areas 122a, 122b, 124a, 124b is typically 0.04" to 0.1 " to create adequate squish without undue heat losses for atypical engine with a bore and stroke in the 3" to 4" range. The clearance is scaled appropriately for larger engines. The engine cylinder views ofFIGS. 11, 1 la, lib, l ie are approximately 2/3 of full scale of atypical IC engine cylinder. In FIGS, lie and l id, where FIG. 1 Id is an exploded view ofthe central part of FIG. 11 c, are shown a centrally located fuel inj ector 130, operable at moderate pressure as in a GDI engine, or at high pressure as in a diesel engine, located between the two valves and spark plugs with central fuel spray cone 131 covering the central combustion region 123. By spraying the fuel in the central region 123, it does not directly impinge on the spark plug electrodes but locates close to them for a better mixed state for ignition. In FIG. 11 d, the squish flow vectors 2 are shown on the left side at the tip of spark plug 118 (representing the piston near TC) which help prevent fuel from settling on the spark plug end and clean the end; the fuel spray is shown on the right side (representing the piston position prior to ignition), being directed by the slight cup 133 in the piston up and to the right as a curved spray 132 moving towards the spark gap 125 b to be ignited. This combination of flow, fuel spray, and ignition location and intensity (high energy), make for a practical system. An advantage ofthe high energy ignition as used in this engine design is that it can sustain flows as high as 20 meters/second (m/s) without spark break-up and can deliver about 100 mJ of spark energy, which allows for very rapid ignition and rapid early flame propagation (which occurs in large part due to the high speed squish and turbulent colliding flows). This allows for ignition timing near TC even for lean and high EGR (or high internal exhaust residual) mixtures, which has several advantages including lower bum temperature and less time available for NOx formation, to minimize NOx emissions. Also, the more recessed exhaust valve 121b allows for more of the partially burnt gas formed at the piston ring regions 126 to be trapped in the combustion volume under the valves than to be exhausted because ofthe long travel path, allowing this partially burnt mixture to be rebumed which minimizes hydrocarbon (HC) emissions. Also, the combination of rapid bum, the larger combustion volume under the exhaust valve 121b, and the large squish zones where the end gas (last part ofthe mixture to bum) is located and subjected to cooling by the reverse squish flow after TC, means that higher compression ratios can be used without engine knock for higher efficiency. Moreover, the flat or slightly cupped piston can be more readily coated with thermal barrier coatings (TBC) to rninimize heat transfer, reduce HC, and increase engine efficiency. FIGS. 12a, 12b, 12c, and 12d are approximately to-scale partial top and side views of a combustion chamber ofFIGS . 11, 11a and l ib detailing and emphasizing certain aspects ofthe combustion chamber for improved operation. Like numerals represent like parts with respect to the earlier figures. FIG. 12a is a partial top view ofthe combustion chamber similar to FIG. 1 la. However, in this case the combustion zone 110 wraps tightly around the outer edges ofthe valves 121a and 121b to have a tight fit with the edge ofthe combustion zone, defining the tight fit regions 141a and 141b, also shown in partial side view FIGS. 12c and 12d. This allows for larger squish zones 122a and 122b on the smaller squish lands which are important in this case not for flow-coupling (produced by the large squish lands 124a and 124b) but for cooling ofthe end gas to permit higher compression ratio operation of 10 to 12 to one compression ratio without knock. The spark plugs 118 and 118a are shown located more towards the larger combustion volume 110a under the exhaust valve versus in-line with the central region 123 in between the two valves, for shorter flame travel and faster burn.. FIGS. 12b and 12c are partial side views ofthe combustion chamber 12a, with FIG. 12b showing conventional type spark plugs 118 and 118a with squish bulk flow (vectors 2) induced by the piston 119b near TC. The side view of FIG. 12c shows views of the two combustion volumes 110a and 110b, approximately scaled to each other, under the exhaust and intake valves respectively. The piston shows an optional combustion zone region 142 cci-fcrniing to the combustion zone in the cylinder head which would result in smaller volumes in the cylinder head. FIG. 12d is a partial view of FIG. 12c showing the intake valve 121a at its approximately fully open position with intake flow vectors 143a emerging from the intake 143 with larger flow (vectors) directed towards the center and right-most region ofthe combustion chamber which can generate tumble flow, useful for flow coupling to the spark gaps at more advanced ignition firing times (shown also with reference to FIGS. 8a and 8b). Likewise, exhaust flow vectors 144a are shown leaving the combustion chamber into the exhaust runner 144, with partially open exhaust valve 121b, with minimum flow (small flow vectors 144a) along the far edge 141b nearest to the piston ring crevice where unburnt hydrocarbons are formed, to help trap them as part of the exhaust residual, as already discussed, with the major exhaust coming from the central region of the combustion zone 11 (large flow vectors 144a). The net result is that more unburnt hydrocar- bons are trapped and higher compression ratios can be used because ofthe squish land et the entire far edge ofthe combustion chamber (where preferably the clearance at TC is approximately 0.04" as is known to those versed in the art). FIG. 13 is a valve tirning diagram depicting a preferred exhaust valve closing (EC) at TC for trapping more exhaust residual, and late intake closing (70° ABC) for higher part-load efficiency and higher, high engine speed output power (higher high speed volumetric efficiency). Intake opening (IO) and exhaust opening (EO) are shown 30° BTC and 30° BBC respectively. FIG. 14 is a design ofthe ignition coil output structure for providing dual ignition firing to the piston ofthe present invention in which the return path for the electrical current is entirely though the spark plugs and not through the engine bloc Like numerals repres ent like parts with respect to the earlier figures . In this preferred embodiment, the coil secondary winding 146 (primary 147) has two floating high voltage outputs connected to spark plug electrodes 10a and 10b which fire to the surface ofthe piston 5 at tips 1 la and 1 lb respectively, as already disclosed . The advantage in this design is that current flows from one electrode (say 10a shown as positive) to the piston surface 1 la and back again to the other electrode 10b at piston surface point 1 lb, so that the spark current, which produces two sparks to the piston surface, does not have to find a return path along the piston/cylinder interface 148 or through other parts connected to the piston. Preferably, one spark gap 149a is smaller than the other (149b) to reduce the required spark firing voltage, or the two plugs have differing capacitances to divide unevenly the output voltage, as is known to those versed in the art. FIG. 15 a is a partial, not to-scale, top view layout of two side-by-side engine cylinders in amulti-cylinder engine with apreferred combustion chamber FIG. 15b ofthe present invention in which the valves are lined up approximately in the length axis ofthe engine and a single overhead cam 150 (broken lines) is shown for actuating the valves (which can also be actuated by other means, i. e. push-rods and rockers, electronically, etc). Like numerals represent like parts with respect to the earher figures. This embodiment is more suitable for swirl generation as dictated by the intake runner 143 which is at the side of a combustion chamber diameter line which produces flow vectors 151 and 152(FIG15b)atthe sparkplug sites (two plugs indicated although plug 118 is preferred if only one plug is used since it will move the spark towards a more central part ofthe combustion zone 110a.). The squish flow vectors are not shown (but will be approximately orthogonal to the swirl vectors). The exhaust runner may be on the opposite side or same side as the intake runner, as indicated in FIG. 15a, it being understood that in an actual engine one or the other would be used (the drawing showing two options, as is also done in FIG. 16a, to conserve drawing space). The combustion zone is indicated as pear shaped with larger diameter zone 110a under the exhaust valve. FIG. 16a is a partial, not to-scale top view layout of two side-by-side engine cylinders in a multi-cylinder engine with apreferred combustion chamber FIG.16b ofthe present invention in which the valves are lined up transverse to the length axis ofthe engine. Like numerals represent like parts with respect to the earher figures . This embodiment is more suitable for tumble flow generation as dictated by the intake runner which is along a combustion chamber diameter line. One ortwo intake valves (121aaand 121ab) are indicated. Two (double) overhead cams 153 and 154 (broken lines) are indicated for separate operation ofthe intake and exhaust valves, which makes it easy to provide independent variable (intake and exhaust) valve timing. In FIG. 16b the version with two intake valves 121aa and 121ab and two intake runners 143a and 143b are shown, which can allow for swirl generation by deactivating one ofthe intake runners, resulting in flow vectors 155 and 156 if valve IN2 ( 121 ab) is deactivated. Squish vectors 2 are also shown in this figure. FIGS. 17a, 17b andl 7c are partial top views of low inductance, high energy ignition coil structures with open magnetic paths suitable for application where greater than 12 volts power is available and designed for having two high voltage towers 160a and 160b suitable for dual ignition. As disclosed elsewhere, this low inductance structure normally has a concentric winding structure with preferably a two layer primary winding 161 and secondary winding 162 made up of typically segmented bobbin winding. Each of the designs have two or more open ends 163 a, 163b (and 163c and 163d for FIG. 17b) to provide a typical low primary winding inductance Lp of 0.3 milliHenry (mH) to 1.0 mH. The coil of FIG. 17a features two side-by-side U-cores 164 and 165 of approximately equal area of each leg placed 180° to each other so that the secondary winding 162 can have its two ends brought out at die two open ends at the high voltage towers for compactness. The high voltage towers 160a and 160b act to also place a clearance between the magnetic fields 166a and 166b and any external metallic surface where the fields might otherwise couple. Likewise for FIG. 17b where two of the four open end magnetic field lines are shown (166a and 166b), FIG. 17c where one of the two open end magnetic field lines are shown. In FIG. 17a preferably the core is made of stacked U-laminations. FIG. 17b features a center leg magnetic core 167 (preferably stacked lamination) and outer magnetic core 168 which can be a tube of wound insulated magnetic tape , e.g. 4 mil silicon iron tape, or two stacked laminations, or other material of different type or shape. Three of four mounting holes 169a, 169b, and 169c are shown. FIG. 17c is a two piece 170a, 170b bobbin type magnetic core, preferably made of stacked laminations, disclosed elsewhere but here designed to have the two high voltage outputs 160a, 160b at the two ends. FIG. 18 is atop partial view of an open E-type low inductance, high energy ignition coil with open magnetic path with some lower turns loosely coupled secondary windings constituting a series of secondary circuit external leakage inductors, usable also in the designs ofFIGS. 17a and 17b, for reducing the end effect voltage overshoot. FIG. 18a is an equivalent circuit ofthe coil ofFIG. 18 used in an inductive ignition circuit. Likenumerals represent like parts with respectto the earher figures. InFIG. 18, the secondary windmg 162 is a segmented windingseparated into bays (10 shown as shaded rectangles), with the lastthree 162abeingfhinner and having fewer turns as well as extending beyond the inner concentric primary winding 161 and having looser magnetic coupling as indicated by the magnetic field lines 170 shown in the left half of the drawing. This is to reduce the bay-to-bay voltage immediately following breakdown ofthe spark gap (gap 15 ofFIG. 18a with gap capacitance 11 a across it). The three windings are chosen by way of example only. In FIG. 18a the last three windings are sho n as three external inductors 171 of total inductance Lend to which can be added an external inductor Lext, as disclosed elsewhere, to reduce the voltage overshoot end-effect immediately following spark breakdown. That is, upon spark breakdown, the high voltage terminal across the spark gap immediately drops to zero volts, and the voltages at the end of the various segments cannot drop as quickly, so that an anomalously high voltage can appear across the secondary bays. In the case where two spark plugs are shown per combustion chamber, a range of options exist for firing them. Under some conditions, e.g. wide-open-throttle (WOT), only one plug may be fired. In some conditions, the plug firing may not be simultaneous. In some conditions, e.g. cold start, the plugs may be multi-fired. Since certain changes may be made in the above combustion chamber designs, in the location and use ofthe ignition, and in the air-fuel mixture flows without departing from the scope ofthe invention herein disclosed, it is intended that all matter contained in the above description, or shown in the accompanying drawings, shall be interpreted in an illustrative and not limiting sense.

Claims: What is claimed is: 1. An internal combustion engine for igniting, combusting, and expanding an air-fuel mixture and producing work by means of a movable piston powered by the combustion-expansion, the engine employing a colliding-flow-coupled-spark-discharge process, CFCSD, comprising a combustion chamber placed and shaped to form a combustion zone mainly within the cylinder head and shaped to form high air-flow regions generated therein by the engine piston motion as it nears the cylinder head on the compression stroke with colliding air flows, in combination with a high energy ignition system with one or more spark plugs located at the edge ofthe combustion zone in the high air-flow regions generated therein by the engine piston motion as it nears the cylinder head on the compression stroke, the system constructed and arranged such that as the piston approaches the cylinder head, one or more ofthe spark plugs are fired to produce sparks which are forced at least radially inwards by the piston induced air-flows towards the center ofthe combustion zone where there is produced intense turbulence, which results in rapid combustion ofthe air-fuel mixture. 2. A CFCSD system as defined in claim 1 wherein the high air-flows are ofthe squish type produced by the piston surface, when near top center (TC) ofthe cylinder, generated by closely spaced cylinder head regions around said combustion zone. 3. An engine ignition system as defined in claim 1 wherein said combustion zone is located mainly under the exhaust valve means. 4. An engine ignition system as defined in claim 1 wherein said combustion zone is located mainly under two centrally located small adjacent valves, the intake and exhaust valve. 5. An engine ignition system as defined in claim 1 wherein said ignition system has flow resistant spark of peak current between 300 and 800 ma. 6. An engine ignition system as defined in claim 1 wherein two spark plugs are used at two opposite edges ofthe combustion zone where high air-flows are generated at the time of ignition firing. 7. An engine ignition system as defined in claim 1 wherein said ignition system is multi-fired during cold start ofthe engine to produce two or more ignition sparks. 8. An engine ignition system as definedin claim 1 wherein the time of ignition firing during cold start ofthe engine is substantially retarded from minimum best timing. 9. An engine ignition system as defined in claim 1 wherein an air-fuel ratio just lean of stoichiometric is used during cold start ofthe engine. 10. An engine ignition system as defined in claim 6 wherein only one of said spark plugs are fired under high engine load conditions with stoichiometric or rich of stoichiometric air-fuel ratios. 11. An engine ignition system as defined in claim 1 further comprising fuel injection means for introducing fuel directly into the combustion zone. 12. An engine ignition system as defined in claim 1 wherein the spark formed from said one or more spark plugs is formed between the center high voltage electrode ofthe spark plug and the piston surface when die piston is near TC. 13. An engine ignition system as defined in claim 4 wherein said combustion zone as proj ected onto die piston face is defined by arcs around and substantially paralleling the outer edges ofthe two valves and joined by essentially straight line barriers defining large squish lands on the two sides of said lines at whose inner edges one or two spark plugs are located to enable high inward radial squish flow to impinge on their sparks formed during ignition firing. 14. An engine ignition system as defined in claim 13 wherein one spark plug is located at said edge ofthe combustion zone and a fuel injector is located at the opposite side ofthe combustion zone. 15. An engine ignition system as defined in claim 13 wherein a fuel injector is placed in the center ofthe combustion zone between the two ports ofthe valves. 16. An engine ignition system as defined in claim 13 wherein two spark plugs are used which form an angle of 15 to 45 degrees with the vertical with their spark gaps in a central portion of the squish flow to direct the sparks radially in and upwards. 17. An engine ignition system as defined in claim 13 wherein the engine compression ratio is between 10 and 12 to one. 18. An engine ignition system as defined in claim 13 wherein said combustion zone has a larger volume under the exhaust valve than the intake valve. 19. An engine ignition system as defined in claim 18 wherein a squish land exists at the entire periphery ofthe combustion chamber to cool the end gas to allow for ahigher engine compression ratio and to increase the path to the exhaust valve and the ease of exit ofthe bum mixture located at the periphery ofthe combustion chamber and piston ring area to trap unburnt hydrocarbons. 20. An engine ignition system as defined in claim 18 wherein said arcs located around the outer edges ofthe two valves are close to the outer valve edges and wherein the intake valve directs its flow in a central direction away from its outer edge to produce tumble type flow which can produce bulk flow at the spark plug site, especially at advanced ignition timing. 21. An engine ignition system as defined in claim 13 wherein said essentially straight line barriers somewhat diverge near the exhaust valve resulting in a larger combustion zone part under and near the exhaust valve relative to the intake valve. 22. An engine ignition system as defined in claim 13 wherein a squish land exists at the entire periphery of the combustion chamber to cool the end gas and allow for a higher engine compression ratio. 23. An engine ignition system as defined in claim 1 wherein the exhaust valve is located essentially at the center ofthe combustion chamber with the combustion zone located mainly underneath it. 24. An engine ignition system as defined in claim 23 wherein two intake valves are located side- by-side between the exhaust valve and periphery ofthe combustion chamber. 25. An engine ignition system as defined in claim 1 wherein the combustion chamber is part of a Miller Cycle engine defined by a late intake valve closing and higher than normal expansion ratio. 26. An engine ignition system as defined in claim 1 wherein the exhaust process is such that higher than normal burnt gas is trapped in the combustion chamber after completion ofthe exhaust process for higher exhaust residual. 27. An engine ignition system as defined in claim 4 wherein the combustion chamber is part of a cam type virtual three stroke engine defined by intake and compression stroke approximately Vi ofthe expansion stroke which has an expansion ratio of 10 to one or greater. 28. An engine ignition system as defined in claim 1 wherein two spark plugs are used in the combustion chamber and a dual output coil is used with at leasttwo open magnetic paths wherein the two high voltage towers are located. 29. An engine ignition system as defined in claim 28 wherein the magnetic core of said coil is made up of two U-cores located at 180° of each other with open ends at opposite ends wherein are located the two high voltage towers. 29. An engine ignition system as defined in claim 27 wherein the magnetic core of said coil is made up of a central I-core over which the primary and secondary windings are wound and an outer magnetic core section, which together constitute a core with the two ends open where the two high voltage towers are located. 30. An engine ignition system as defined in claim 1 wherein two spark plugs are used per combustion chamber which fire to the piston and use a dual output coil which allows for the spark current to complete the electrical circuit entirely within the coil dual output secondary winding. 31. An engine ignition system as defined in claim 1 wherein the ignition system used is an inductive ignition with low primary inductance and with at least one open magnetic path end where the secondary winding,, comprised of segmented bays, extends so that at least part or all ofthe last bay is poorly coupled to the magnetic field produced by the primary winding. 32. An engine ignition system as defined in claim 1 wherein the engine compression ratio is between 10 and 12 to one. 33. An improved ignition-combustion system for internal combustion engines comprising: (a) means defining a compact combustion chamber zone mainly in the engine cylinder head with at least approximately 50% of it under the exhaust means, (b) means for forming large air-squish zones formed at the edge ofthe combustion zone which produce colliding squish flows with high turbulence at the center ofthe combustion zone, (c) one or more spark plug means located at the edge ofthe combustion zone with at least one spark plug means located within the high squish zone, (d) the ignition employing high energy flow-resistant ignition sparks which move under the influence ofthe squish flow towards the central turbulence region as the pistonnears engine top center (TC) and ignition firing occurs near TC when squish is high, the squish flow further reinforced by other bulk flow at the spark plug site when ignition occurs substantially earlier of TC when squish flow is relatively low, to produce rapid and complete burning of lean and high EGR mixtures whereby high engine efficiency and lowest emissions are achieved. 34. An ignition system as defined in claim 31 wherein said other bulk flow is swirl or tumble flow. 35. An improved ignition-combustion system for internal combustion engines with high air-flows in the combustion chamber during the time of ignition employing a high energy ignition system ofthe low inductance type of primary inductance below 2 mH with coils constructed with one or more open ends at least at the high voltage ends with some ofthe high voltage end secondary winding extending out so as to be coupled poorly by the magnetic field generated by the primary winding such that said extended high voltage end secondary winding represents series inductors in the secondary circuit to reduce the end effect over- voltage. 36. An ignition system as defined in claim 35 wherein said secondary winding is a segmented winding with at least few bays having fewer turn that the previous bays and with part or all of said last winding bays extending beyond the primary winding at the magnetic core open end.

Inventors: WARD MICHAEL A V, WARD, MICHAEL, A., V.
Applicants: COMB ELECTROMAGNETICS INC, WARD MICHAEL A V, COMBUSTION ELECTROMAGNETICS, INC., WARD, MICHAEL, A., V.
Application year: 2000
Main classification: F02B23
IPC codes: F02B17, F02B23, F02B31