Patent Abstract:
An internal combustion engine uses separate, tubular and hollow reciprocating sleeve valves that open and close intake and exhaust passageways for improved aspiration. The sliding sleeve valves are disposed within sleeves horizontally disposed within a modified head secured above the combustion chamber. The valves are driven in a path normal to the engine pistons by an independent crankshaft that is rotated through an external pulley driven by the engine crankshaft. Fluid flow occurs through the valve interior and through ports dynamically positioned above the compression cylinder, proximate aligned sleeve and head ports. Sleeve ports are separated by bridges that maintain valve rings in compression during reciprocation to prevent damage. Each valve body has a reduced diameter midsection forming a relief annulus that distributes shearing pressures about the circumference of the valve. High pressure gas is confined between axially spaced apart, stepped sealing rings that prevent gases from flowing axially about the valve exterior.

Full Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This utility patent application is based upon, and claims the filing date of, prior U.S. Provisional application entitled “Sliding Valve Aspiration Engine,” Ser. No. 61/135,267, filed Jul. 18, 2008, by inventor Gary W. Cotton. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates generally to sleeve valve systems for aspirating internal combustion engines, and to internal combustion engines with tubular sliding valves for enhanced aspiration. More particularly, the present invention relates to reciprocating sleeve valve systems and engines equipped therewith of the general type classified in United States Patent Class 123, Subclasses 84, 188.4, and 188.5. 
     2. Description of the Related Art 
     A variety of aspiration schemes are recognized in the internal combustion motor arts. In a typical four-cycle firing sequence, gases are first inputted and then withdrawn from the combustion chamber of each cylinder interior during reciprocating piston movements caused by the crankshaft. Gas pathways must be opened and closed during a typical cycle. During the intake stroke, for example, an air/fuel mixture is suctioned through an open intake passageway into the combustion chamber as the piston is drawn downwardly within the cylinder. The intake passageway is typically opened and closed by some form of reciprocating valve mechanism that is ultimately driven by mechanical interconnection to the crankshaft. The combustion chamber must be sealed during the following compression and power strokes, and the valve mechanisms must be closed to block the ports. During the following exhaust stroke, exhaust ports must be opened to discharge spent gases from the combustion chamber. 
     Spring-biased poppet valves are the most common form of internal combustion engine valve. Typically, poppet valves associated with the intake and exhaust passageways are seated within the cylinder head above the combustion chamber proximate the cylinder and piston. Typical reciprocating poppet valves are spring biased, assuming a normally closed position when not deflected. In a typical arrangement, the bias spring coaxially surrounds the valve stem to maintain the integral valve within the matingly-configured valve seat. Poppet valves are typically opened by mechanical deflection from valve train apparatus driven by camshafts. Typical overhead-valve motor designs include rocker arms comprising reciprocating levers driven by push rods in contact with camshaft lobes. When the camshaft lobe deflects a pushrod to raise one end of the rocker arm, the opposite arm end pivots downwardly and opens the valve. When the camshaft rotates further, the rocker arm relaxes and spring pressure closes the valve. With overhead-cam designs camshafts are disposed over the valves above the head, and valve deflection is accomplished without push rods or rocker arms. Overhead camshafts push directly on the valve stem through cam followers or tappets. Some V-configured engines use twin overhead camshafts, one for each head. Some enhanced DOHC designs use two camshafts in each head, one for the intake valves and one for the exhaust valves. The camshafts are driven by the crankshaft through gears, chains, or belts. 
     Despite the overwhelming commercial success of poppet-valve designs, there are numerous deficiencies and disadvantages associated with poppet valves. Although poppet valve designs provide manufacturing advantages and cost savings, substantial spring pressure must be repeatedly overcome to properly open the valves. Spring pressure results in considerable drag and friction which increases fuel consumption and limits engine RPM. Poppet valve heads are left within the fluid flow passageway, despite camshaft deflection, and the resulting obstruction in the gas flow pathway promotes inefficiency. For example, back pressure is increased by the valve mass obstructing fluid flow, which contributes to turbulence. Poppet valves are exposed to high combustion chamber temperatures, particularly during the exhaust stroke, that can promote deformation and wear. Thermal expansion of exhaust valves, for example, can interfere with proper valve seating and subsequent sealing, which can decrease combustion performance. 
     Many of these disadvantages are amplified in high-horsepower or “high R.P.M.” applications. Valve deflection in high power applications is often extreme, increasing the amplitude of valve defection or travel. Damaging valve-to-piston contact can result. As a means of attenuating the latter factor, some pistons are designed with valve clearance regions, but these piston surface irregularities can deleteriously affect the combustion charge and fluid flow through the combustion chamber. Another problem is that the applied drive forces experienced by the valves are asymmetric. The extreme forcing pressure applied by the camshaft to open the valves, for example, is not as uniform as the spring closing pressure. Disharmony between the opening and closing forces contributes to valve lash and concomitant timing problems that interfere with power generation and limit engine R.P.M. Of course, in high power systems involving four or more valves per cylinder, the problems and disadvantages with poppet valve engines are increased proportionally. 
     So-called “rotary valves” have been proposed for replacing reciprocal poppet valves. Typical rotary valve designs include an elongated tube or cylinder machined with a plurality of gas flow passageways that admit or pass gases. The rotary valves are not reciprocated; the are rotated about their axis to expose passages defined in them in directions normal to their longitudinal axis. Rotary valves must be timed properly to dynamically align their internal passageways with the fluid flow paths of the engine during operation. When rotated to a closing position, the rotary valve passageways are radially displaced, obstructing the normal flow pathways and sealing the engine for firing or compression strokes. 
     One advantage espoused by rotary valve proponents is the relative simplicity of the design. Further, rotary valves do not penetrate or extend into the cylinder, avoiding potential mechanical contact with the piston, and minimizing fluid flow obstructions. However, the biggest problem with rotary valves relates to ineffective sealing. Although much activity and research has been directed to rotary valve sealing designs, commercially feasible systems have not been perfected. Rotary systems provide inefficient cylinder sealing, lessening firing efficiency, and reducing compression pressure because of leakage. Further, rapid wear of such systems increases the aforementioned problems. 
     Sliding valves of many configurations are also known in the art. Typical slide valves may be hollow and tubular, or planar, or cylindrical. They are reciprocated within a tubular valve seat region proximate the combustion chamber to alternately open and then close the intake and exhaust passageways. Like rotary valves, sliding valve designs have hitherto been difficult to seal effectively, with predictable negative results. 
     U.S. Pat. No. 2,080,126 issued May 11, 1937 to Gibson shows a sliding valve arrangement involving a tubular valve driven by a secondary crankshaft. Its reciprocating axis is parallel to the axis of piston deflection. Ports arranged at the side of the piston are alternately opened and closed by piston movements, and gases are conducted through and around portions of the piston exterior. 
     A similar arrangement is seen in U.S. Pat. No. 1,995,307 issued Mar. 26, 1935, and U.S. Pat. No. 2,201,292, issued May 21, 1940, both to Hickey. The latter patents show designs that aspirate a single working cylinder with a pair of tubular, reciprocating valves that are mounted on either side of the piston and driven by secondary crankshafts. The aspirating valves are forcibly reciprocated between port blocking and port aligning positions. The valves are aligned at an angle slightly off of parallel with the axis of the cylinder. 
     Other examples of engines with tubular, reciprocating slide valves that move in a direction generally parallel with the drive piston axis are provided by U.S. Pat. Nos. 1,069,794; 1,142,949; 1,777,792; 1,794,256; 1,855,634; 1,856,348; 1,890,976; 1,905,140; 1,942,648; 2,160,000; and 2,164,522 that are largely cumulative. 
     Hickey U.S. Pat. No. 2,302,442 issued Nov. 17, 1942 shows a tubular, reciprocating sliding valve disposed atop a piston head. The valve slides in an axis generally perpendicular to the axis of the lower drive piston. 
     U.S. Pat. No. 5,694,890 issued to Yazdi on Dec. 9, 1997 and entitled “Internal Combustion Engine With Sliding Valves” discloses an internal combustion engine aspirated by slidable valves. Tapered, horizontally disposed valve seats are defined near inlet and exhaust ports at the top of the combustion chambers. The slidable valves are tapered to conform to the valve seats. Valve movement is caused by a crankshaft driving a rocker arm that is oriented substantially orthogonal to the rod, whereby crankshaft rotation is translated into horizontal, sliding movements of the planar valves, which reciprocate in a direction normal or transverse to the axis of the piston. 
     U.S. Pat. No. 7,263,963 issued to Price on Sep. 4, 2007 and entitled “Valve Apparatus For An Internal Combustion Engine” discloses a cylinder head with a cam-driven valve slidably disposed within a valve pocket. The valve, which is displaceable along its longitudinal axis has a tapered portion defining multiple fluid flow passageways. The valve is displaced by cam rotation between a configurations passing gases through the passageways and a configuration wherein the valve flow passageways are closed. 
     BRIEF SUMMARY OF THE INVENTION 
     This invention provides an improved sliding valve system for aspirating internal combustion engines, and engines equipped therewith. The system employs tubular, reciprocating sliding valves disposed within sleeves defined within the head secured above the motor&#39;s reciprocating pistons. The valves are driven by an independent crankshaft that is exteriorly driven through a pulley. 
     The sliding valves are positioned within suitable exhaust and intake tunnels in the head. Preferably sleeves are concentrically disposed around the valves and concentrically fitted within the tunnels. Fluid flow through the valves results through ports defined in the body of the tubular slide valves that are aligned with similar ports in their sleeve, that are in turn aligned with ports dynamically positioned above the compression or combustion region of the cylinder located below the head. Gas pressure develops shearing forces on valve sides. Gases are routed through the tubular interior of the sliding intake valve or valves during intake strokes, and exhaust gases are likewise forced out of the combustion cylinder through the interior of the exhaust valve or valves during exhaust strokes. Pressured gases traveling longitudinally through the valve interior passageways are inputted or outputted through lateral valve ports in fluid flow communication with the internal valve passageways. 
     Rather than pressuring faces of the valves in a direction normal to valve travel, exhaust and intake gas forces are directed against sides of the valves. To minimize potentially detrimental forces applied across the valves during, for example, the critical exhaust stroke, the valve body includes at least one reduced diameter portion forming a relief annulus within the valve chamber that distributes potential shearing pressure about the circumference of the valve. High pressure gas is confined between axially spaced apart sealing rings that prevent gases from flowing axially about the valve exterior. 
     All intake and exhaust gas flow is thus confined within the tubular interior of the valves. As a result, gas pressure does not develop a substantial resistive force upon leading surfaces of the valve in a direction coincident with the direction of valve travel. Instead gas pressure that might otherwise resist valve travel, and add to friction, is applied as a shear force, and pressure is evenly distributed in the relief annulus. Gas flow is distributed through the valve interior rather than around it, and friction is substantially reduced. 
     Importantly, the port sizes are maximized for efficient breathing. However, in the past, large sliding valve ports have contributed to inefficiency, reduced sealing, and premature valve failure. In the present design, the slide-valve sleeves are provided with a unique connecting bridge that traverses the port area, aligned with the direction of sliding valve travel. When the valves slidably reciprocate through this region, their sealing rings are supported tangentially by the bridges, to maintain ring integrity. 
     Thus a basic object of my invention is to provide a highly efficient aspiration or valve system for internal combustion engines, particularly four-cycle designs. 
     A related object is to provide an improved four cycle, internal combustion engine. 
     A related object is to improve combustion efficiency within an internal combustion engine. It is a feature of our invention that its advantageous overhead valve geometry and the reduction of valve-train parts needed for the invention increase overall efficiency. 
     Another important object is to preserve the sealing integrity of sliding valves. One important feature of the invention in this regard is that the head ports are provided with bridges that support the valve sealing rings during motion. 
     Another basic object is to provide a valve system for internal combustion engines that provides an enhanced power stroke. In other words, it is a feature of this invention that a higher proportion of the total 720 degrees of crankshaft rotation during typical four cycle operation occurs during the power stroke. 
     Another important object is to provide a sliding valve system of the character described that does not affect combustion chamber volume during operation. Important features of my invention are the fact that chamber expansion during valve displacement is avoided, and that the porting path does not consume the operational compression volume. 
     A related object is to provide a valve system of the character described wherein the valve structure does not enter the combustion chambers. 
     Another object is to provide a valve deflection system that applies force symmetrically, to minimize valve lash and allow higher engine speeds. 
     Yet another basic object is to minimize friction. It is a feature of my invention that spring-biased poppet valves and the typical frictional cam shafts and associate linkages such as rocker arms used to reciprocate poppet valves are avoided. 
     A still further object is to provide a valve system of the character described that is driven externally by a belt, so that efficiency is increased and complexity is reduced. 
     Another important object is to avoid so-called split-lift” applications used in the prior art for aspirating motors. 
     These and other objects and advantages of the present invention, along with features of novelty appurtenant thereto, will appear or become apparent in the course of the following descriptive sections. 
    
    
     
       BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS 
       In the following drawings, which form a part of the specification and which are to be construed in conjunction therewith, and in which like reference numerals have been employed throughout wherever possible to indicate like parts in the various views: 
         FIG. 1  is a fragmentary isometric view of a one-cylinder internal combustion engine constructed in accordance with the best mode of the invention known at this time; 
         FIG. 2  is an enlarged, fragmentary, plan view of the engine taken generally from a position to the right of  FIG. 1  and looking left, with portions thereof broken away or shown in section for clarity; 
         FIG. 3  is an enlarged, fragmentary sectional view taken generally along line  3 - 3  of  FIG. 2 ; 
         FIG. 3A  is a greatly enlarged, fragmentary view of circled region  3 A in  FIG. 3 ; 
         FIG. 4  is an enlarged, fragmentary, isometric view of the preferred cylinder head assembly, with portions thereof broken away or shown in section for clarity or omitted for brevity; 
         FIG. 4A  is a greatly enlarged, fragmentary view of circled region  4 A in  FIG. 4 ; 
         FIG. 5  is an enlarged, partially exploded fragmentary isometric view of the cylinder head assembly of  FIG. 4 , with a sliding valve removed from its sleeve, and with portions thereof broken away or shown in section for clarity; 
         FIG. 6  is an enlarged, fragmentary isometric view taken generally from circled region “ 6 ” in  FIG. 5 ; 
         FIG. 7  is an enlarged bottom isometric view of the preferred cylinder head; 
         FIG. 8  is an enlarged isometric view of a preferred spool valve, with portions thereof broken away or shown in section for clarity; 
         FIG. 9  is a side elevational view of a preferred spool valve; 
         FIG. 10  is an end elevational view of the spool valve of  FIG. 9 , looking generally in the direction of arrows  10 - 10 ; 
         FIG. 10A  is a longitudinal sectional view of a preferred spool valve, derived generally in the direction of arrows  10 A- 10 A in  FIG. 10 ; 
         FIG. 11  is an enlarged top plan view of the preferred cylinder head, with phantom lines illustrating various internal parts, and with portions broken away or shown in section for clarity; 
         FIG. 12  is an enlarged, fragmentary diagrammatic view showing the basic arrangement of the engine power cylinder, the head, the overhead spool exhaust valve, and the exhaust valve sleeve; 
         FIGS. 13-15  are diagrammatic views of progressive intake spool valve movements during the intake stroke as the power crankshaft rotates; 
         FIG. 16  is a diagrammatic view showing the intake spool valve position when the spark plug fires at the beginning of the power stroke; 
         FIG. 17  is a diagrammatic view showing the intake spool valve position at the bottom of the power stroke; 
         FIG. 18  is a diagrammatic view showing the intake spool valve position at the end of the exhaust stroke; 
         FIG. 19  is a diagrammatic view showing the exhaust spool valve position at the start of the exhaust stroke; 
         FIG. 20  is a diagrammatic view showing the fully open exhaust spool valve position at 251 degrees of engine crankshaft angle; 
         FIG. 21  is a diagrammatic view showing the closing exhaust valve at the beginning of the intake stroke at 222 degrees of crankshaft angle; 
         FIG. 22  is a diagrammatic view showing the fully closed exhaust valve at the bottom of the intake stroke at 180 degrees of crankshaft angle; 
         FIG. 23  is a diagrammatic view showing the closed exhaust valve 90 degrees into the compression stroke; 
         FIG. 24  is a diagrammatic view showing the closed exhaust valve at zero degrees TDC; 
         FIG. 25  is a longitudinal diagrammatic view of the preferred secondary crankshaft that operates the intake and exhaust spool valves and moves them between positions illustrated in  FIGS. 13-24 ; 
         FIGS. 26-28  are sectional views taken respectively along lines  26 - 26 ,  27 - 27 , and  28 - 28  of  FIG. 25 ; 
         FIG. 29  is an isometric view of a preferred spool valve sleeve, with portions broken away for clarity; 
         FIG. 30  is a bottom plan view of the sleeve of  FIG. 29 ; 
         FIG. 31  is a side elevational view of the sleeve of  FIG. 29 ; 
         FIG. 32  is an end elevational view of the sleeve of  FIG. 29 ; 
         FIG. 33  is an enlarged, side elevational view of a preferred sealing ring used with the sliding valves; 
         FIG. 34  is an enlarged, plan view of a preferred sealing ring used with the sliding valves; and, 
         FIG. 35  is an enlarged, fragmentary plan view of circled region  35  in  FIG. 33 . 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     With initial reference directed to  FIGS. 1-3 ,  3 A,  4 ,  4 A, and  5  of the appended drawings, a basic single-cylinder, four-cycle internal combustion engine equipped with the aspiration system constructed generally in accordance with the best mode of the invention has been generally designated by the reference numeral  10 . It should be understood that the aspiration system as herein described is suitable for use with engines equipped with multiple cylinders, arrayed in the popular V-configuration or other configurations. The engine  10  has a rigid block  11  housing a primary crankshaft  12  ( FIG. 3 ) of conventional construction that drives a reciprocating power piston  14  ( FIG. 3 ) with a conventional connecting rod  16 . The basic engine illustrated comprises a Honda thirteen-horsepower motor, which is modified as hereinafter described. 
     The standard combustion power piston  14  reciprocates within a cylinder  18  ( FIG. 3 ) that is externally air-cooled with multiple external heat dissipation fins  20  ( FIG. 1 ) proximate the engine deck  13 . The basic construction of the conventional piston  14  and its accessories is substantially conventional and is not critical to practice of the invention. The instant sliding valve system is disposed within a head, generally indicated by the reference numeral  22  (i.e.,  FIGS. 4 ,  5 ,  7 ,  11 ), that mounts conventionally above the engine deck  13  above the conventional piston  14  and cylinder  18  described previously. The stroke of power piston  14  moves it upwardly and downwardly in a direction substantially perpendicular to head  11 . For purposes of this invention, the term “head” shall generally designate that region of an internal combustion engine enclosing the combustion chambers, above the pistons. Such a head may be a conventional separate part bolted atop the engine, or in some cases the “head” may be integral with the engine block in a single casting that is thereafter appropriately machined. 
     With additional reference directed primarily now to  FIGS. 4-11 , head  22  houses a pair of tubular, sliding spool valves  24 ,  25  ( FIGS. 8-10 ) that aspirate the cylinder  18 . Based upon experiments so far, the tubular exhaust valve  24  and the tubular intake valve  25  are made from titanium in the best mode. While those skilled in the art will recognize that several alloys of titanium and/or titanium steel are available, my experiments have yet to reveal the ideal composition of these critical valves. Ordinary steel compositions however, result in heat damage and premature wear and failure. Furthermore, as illustrated in  FIG. 5 , for example, the sliding valves  24 ,  25  are mounted in appropriately ported sleeves  27  that fit into the cylinder head and line up with the sliding valve ports and appropriate ports in the head. However, experiments with the engine as depicted with sleeveless valves have shown the design to be rugged and dependable so far. 
     A drive pulley  26  ( FIG. 1 ) driven by conventional internal crankshaft  12  ( FIG. 3 ) is connected via drive belt  28  to a valve pulley  30  that drives the slide valve crankshaft  32  housed within head  22 . Crankshaft  32 , best seen in  FIG. 25  discussed hereinafter, is mounted perpendicularly relative to sliding valves  24 ,  25  (i.e.,  FIGS. 7 ,  11 ). It extends across and through compartmentalized crankshaft mounting region  34  ( FIG. 5 ) across the top (i.e., as viewed in  FIGS. 4 ,  5 ) of the head  22 . Region  34  contains liquid oil for lubricating the crankshaft and the slide valves to be described. Region  34  is normally covered by shroud  35  ( FIG. 3 ). The crankshaft exhaust journal  38  and the crankshaft intake valve journal  40  (i.e.,  FIG. 25 ) of crankshaft  32  support connecting rods  42 ,  44  that respectively operate exhaust slide valve  24 , and intake slide valve  25 . Aligned and integral crankshaft portions  39 ,  41 ,  43  (i.e.,  FIG. 25 ) are rotatably constrained within conventional saddles  45  within mounting region  34  (i.e.  FIG. 4 ,  5 ) and mounted with conventional bearing assemblies  46  ( FIG. 2 ) as known in the art. In the best mode it is proposed that the counterweight sections  109 ,  110 ,  111 , and  112  of the crankshaft ( FIG. 25 ) be drilled appropriately for crankshaft balancing. Preferably the rotating and reciprocating aspiration slide valve assembly may thus be “balanced” and “tuned” for optimal aspiration performance. 
     The crankshaft bearing assemblies  46  are bolted within crankshaft region  34  to mount the slide valve crankshaft  32  over the saddles  45  are secured with a plurality of bolts  48 . As best seen in FIGS.  4 , 5  and  7 , head  22  includes a plurality of spaced apart mounting orifices  50  through which head bolts  52  ( FIG. 11 ) extend when mounting the head  22  to the deck  13 . 
     The intake spool valve  25  (i.e.,  FIG. 11 ) is slidably received within a sleeve  27 B disposed within head tunnel  55  ( FIGS. 4 ,  11 ), that is spaced apart from and parallel with exhaust tunnel  54  and sleeve  27 . Tunnels  54  and  55  are oriented generally perpendicularly to the stroke of the power piston  14 . Exhaust spool valve  24  slidably reciprocates within sleeve  27  concentrically disposed within tunnel  54 . Sleeves  27 ,  27 B ( FIGS. 5 ,  29 - 32 ) require ports aligned with head ports and valve described hereinafter, as appreciated by those skilled in the art. An air-fuel mixture is drawn into intake valve tunnel  55  from a conventional carburetor  29  ( FIG. 2 ) mounted with screws received within orifices  59  ( FIG. 4 ). Alternatively the invention may be used with fuel injection systems. 
     As best viewed in  FIGS. 29-32 , each sleeve  27  is elongated and tubular. Each has a pair of spaced apart open ends  31  defining opposite ends of an elongated cylindrical passageway in which the sliding valves  24  and/or  25  are inserted. A pair of ports  68 A are separated by a bridge  69 A ( FIG. 29 ) that maintains pressure on the sliding valve rings during operation. While both sleeves are identical in dimensions and geometry, the exhaust sleeve should be of a more expensive heat resistant alloy. It is preferred that the exhaust sleeve be made of Steelite or Nickalloy heat resistant titanium steel alloy. 
     This invention requires maximal air flow quickly. In other words, it is preferred that the carburetor  29  have a relatively large throat with a relatively short venturi. In the model depicted in the drawings, which has been thoroughly tested, a Honda 350 cc. “dirt bike” motorcycle carburetor is preferred. 
     Exhaust valve  24  is slidably constrained within its sleeve  27  in tubular tunnel  54  ( FIGS. 5 ,  7 ,  11 ). The exhaust header  57  ( FIG. 1 ) is preferably screw-mounted upon the head&#39;s end surface  58  ( FIGS. 4 ,  7 ) with suitable screws that penetrate orifices  60 . Head cooling is encouraged by fin areas  36  ( FIG. 5 ). 
     As best seen in  FIG. 7 , the circular combustion chamber  62  includes a central, threaded spark plug passageway  64  that is spaced between intake ports, collectively numbered  66 , and exhaust ports, collectively numbered  68  ( FIG. 7 ). A conventional spark plug  70  (i.e.,  FIGS. 1 ,  11 ) is threadably mated to passageway  64 , with its electrodes positioned and centered within combustion chamber  62 . 
     As seen in  FIGS. 29-30 , for example, adjacent sleeve ports  68 A are separated from one another by a central bridge  69 A. Similarly intake ports  66  in the head ( FIG. 7 ) built into the combustion chamber may be separated with a bridge  67  that is integral with the head  22 . Similarly, a rigid, centered bridge  69  in the head separates the twin exhaust ports  68  ( FIGS. 6 ,  7 ). These ports in the head must align with the valve sleeve ports  68 A seen in  FIGS. 29-32 . 
     As best seen in  FIG. 6 , each head exhaust port  68  aligns with sleeve port  68 A. The composite ports have smooth, downwardly inclined sidewalls  74 ,  75  that are polished for maximal fluid flow. These walls communicate with a lower orifice  73  in the head that opens to the combustion chamber  62 . The intake ports  66  (i.e.,  FIG. 7 ) are similarly configured. Importantly, it is desired that corner ridges of the structure be radiused for maximum fluid flow, as illustrated by gently radiused corner regions 
     Importantly, rigid, transverse bridges  69 A are integrally formed in the sleeve port regions and bisect these regions into twin, side by side orifices  68 A ( FIG. 29 ). The head is similarly ported. In  FIG. 7 , for example, there are two pairs of ports  66  and  68  respectively separated by bridges  67 ,  69 . Sleeve  69 A bear against critical sealing rings associated with the sliding valves  24  and  25 , as discussed below. By pressuring the sealing rings during valve travel, deformation of the critical sealing rings in the region of the various exhaust ports  68  and intake ports  66  is prevented. As sealing of the tubular slide valves  24 ,  25  is critical to the invention, bridges  67  and  69  are vital to the best mode of the invention. 
     With joint reference directed now primarily to  FIGS. 8-12  and  10 A, valves  24  and  25  are structurally virtually identical, so only exhaust valve  24  will be detailed. However, it is thought that the exhaust valve  24  requires a more heat resistance, so a premium grade of titanium alloy steel is preferred. 
     Each valve  24 ,  25  is elongated, substantially tubular, and multi-sectioned. An open connecting rod section  80  enables connection to the connecting rod  42  ( FIG. 12 ). The rod end  42  extends into the interior  82  of section  80  and is journalled by wrist pin  85  ( FIG. 3 ) and is conventionally secured between wrist pin orifices  84  ( FIGS. 9 ,  10 A). Importantly, section  80  ends in a closed interior wall  87  that separates region  82  and the connecting rod structure from the rest of the tubular interior  89  ( FIG. 10A ) of the valve  24 . The open end of the interior passageway  89  within each valve directly communicates through tubular tunnels  54 , or  55  ( FIG. 4 ) for aspiration fluid flow. The exterior of valve rod section  80  ( FIGS. 9 ,  10 A) is preferably cross hatched by machining to promote oil flow and distribution. 
     In the best mode each valve has three pairs of external ring grooves to seat suitable sealing rings. For example, a pair of concentric and parallel ring grooves  91  separate valve rod section  80  from port section  94  ( FIG. 9 ). Ring grooves  92  separate port section  94  from adjacent midsection  96 . Similarly, ring grooves  93  separate midsection  96  from open section  98 .  FIG. 8  shows that each pair of ring grooves  91 ,  92  and/or  93  seats pairs of spaced apart, concentric sealing rings  100 A,  100 B and  100 C respectively, that are externally, coaxially mounted about the valve exterior. Since each valve rod section  80  is in fluid flow communication with head region  34  that contains lubricating oil, rings  100 A are oil rings. It will be recognized by those skilled in the art that when the valves  24  or  25  are fitted within their sleeves  27 , (i.e.,  FIG. 4 ) the rings  100 A,  100 B, or  100 C will seat within ring grooves  91 ,  92  or  93  (i.e.,  FIG. 9 ) and the exterior of the rings will be flush with the cylindrical outside body of the valves  24 ,  25 , touching the interior surfaces of the captivating sleeves  27 . 
     Each sealing ring  100 A,  100 B,  100 C is preferably made of heat treated and heat resistant nickel alloy steel. As best seen in  FIGS. 33-35 , the compressively touching ends of the rings are stepped in the best mode to form an overlapped intersection  113  that forms an improved pressure seal. Preferably, each end of a given ring is configured in the overlapping or stepped configuration of  FIG. 35 , where abutting ring ends comprise a notched region  115  and a bordering, elongated tabbed region  116 . The tabbed regions  116  are variably spaced apart from notched regions  115 , with end gaps  117  therebetween. The parallel, spaced apart ring end gaps  117  allow for thermal expansion and contraction of the rings during operation. However, a sealing gap  118 , which is perpendicular to gaps  117 , is defined between mutually aligned and abutting tabbed regions  116 . Gap  118  is much smaller than indicated, and provides a seal, as end regions  116  abut in operation, and seal the gaps for compression. At the same time gaps  117  allow for normal thermal expansion and contraction. 
     Importantly, the valve port section  94  ( FIGS. 8 ,  9 ) includes an enlarged, arcuate cutout  102  functioning as an aspiration port (i.e., either exhaust or intake). Port  102  radially extends about approximately 30-40 percent of the radial periphery of the valve. A gently radiused arch  103  above port  102  ( FIGS. 8 ,  10 A) leads to the smoothly configured, generally cylindrical passageway  89  that leads to the exterior of the valve. Passageway  89  ( FIG. 10A ) comprises tubular interior passageway walls  104 , terminating in gently radiused, flared lips  106  ( FIG. 10A ) at the valve end that maximize fluid flow. Aspiration occurs when valve ports  102  are aligned with sleeve ports  68 A ( FIG. 32 ) which are in turn aligned with head port pairs  66  or  68  ( FIG. 7 ), in response to timed, reciprocal movements caused by the valve crankshaft  32  previously described. Thus when port  102  ( FIGS. 3 ,  9 ) of the exhaust valve  24  overlies sleeve ports  68 A ( FIG. 32 ) and head ports  68  ( FIG. 7 ), hot exhaust gases may be vented away from the combustion chamber  62  and lower cylinder  18  in response to upward movement of the power piston  14  towards top-dead-center. At this time exhaust gases are vented to the left (as viewed in  FIG. 9 ) through port  102 , along the valve interior passageway  89  ( FIG. 8 ) and through head tunnel  54  ( FIG. 7 ) and out header  57  ( FIGS. 1 ,  3 ). Similarly, during the intake stroke, air and raw fuel is drawn through carburetor  29  into the head  22  through tunnel  55  ( FIG. 7 ), and into the chamber  89  in the intake valve  25 , through its port  102  and into the cylinder combustion region through head ports  66  ( FIG. 7 ) and aligned sleeve ports  68 A. 
     Importantly, as slide valves  24 ,  25  reciprocate, their multiple sealing rings  100  are prevented from deformation while traversing sleeve ports  68 A by the bridges  69 A (i.e.,  FIG. 32 ). Further valve deformation is prevented by the downsized diameter of valve midsections  96  (i.e.,  FIG. 8 ). Referencing  FIG. 9 , the arrow  105  indicates the outside diameter of the majority of the length of valve  24 . Sections  80 ,  94 , and  98  are all of this relatively larger diameter. Valve midsection  96  however, has a reduced diameter indicated by the arrow  107  ( FIG. 9 ). When the valves  24 ,  25  are positioned to “block” the various ports, midsection  96  is positioned over them. Thus a cylindrical or annular region  101  ( FIGS. 3 ,  3 A,  4  and  4 A) defined radially around the external periphery of valve midsection  96  between the surrounding tunnels  54  or  55 , and axially defined between the rings  100  on opposite ends of valve midsection  96 , will be in fluid flow communication with the combustion chamber  62 . Annulus  101  thus distributes potential shearing pressure about the circumference of the valve when the ports are blocked during various valve stroke positions to reduce damage. During the power stroke, for example, the shock from rising gas pressure will be uniformly distributed about the radial periphery of valve midsection  96  within annulus  101 , equalizing forces that might otherwise deform the valve. 
     OPERATION 
     In  FIG. 13  intake valve  25  has started to open at the beginning of the intake stroke. In  FIG. 14  the intake valve  25  is now open at approximately 108 degrees BTDC. 
       FIG. 15  shows the intake valve  25  closing at the end of the intake stroke. Full closure of valve  25  is indicated in  FIG. 16  at the beginning of the power stroke. 
       FIG. 17  shows the bottom of the power stroke, with the intake valve  25  fully closed. In  FIG. 18  at the end of the exhaust stroke the intake valve  25  is seen starting to open. 
     The exhaust valve  24  is seen in  FIG. 19  at the start of the exhaust stroke. In  FIG. 19 , the plug and cylinder have fired, and at 108 degrees ATDC the exhaust valve  24  starts to open. In  FIG. 20  the exhaust valve  24  is completely open, with 251 degrees crankshaft angle. 
     At the beginning of the intake stroke in  FIG. 21  the exhaust valve  24  begins to close, at approximately 222 degrees. The bottom of the intake stroke is seen in  FIG. 22 , at which time the exhaust valve  24  is fully “closed,” and the reduced diameter midsection  96  is positioned over the exhaust ports  68 . 
     In  FIG. 23  the exhaust valve  24  is completely open, 90 degrees into the compression stroke. In the positions of  FIG. 24  the plug fires, and the exhaust valve  24  is completely closed at zero degrees TDC. 
     In  FIGS. 25-28  the configuration and position of the crankshaft  32  is illustrated. The exhaust valve journal  40  and the intake journal  38  are seen in critical rotational positions. 
     EXAMPLE 
     Dyno Test Chart—December, 2008 
     
       
         
               
               
               
               
             
           
               
                   
                   
               
               
                   
                   
                 FACTORY 
                   
               
               
                   
                 LOW LOAD 
                 ENGINE 
                 G1 ENGINE 
               
               
                   
                   
               
             
             
               
                   
                 Load % 
                 33% 
                 33% 
               
               
                   
                 RPM 
                 2900 
                 2900 
               
               
                   
                 Run Time 
                 1:30 minutes 
                 1:30 minutes 
               
               
                   
                 lb-ft Torque 
                 7.5 
                 7.5 
               
               
                   
                 Brake Horsepower 
                 4.1 
                 4.1 
               
               
                   
                 Fuel Usage - Milliliters 
                 12.07 
                 10.86 
               
               
                   
                 Nitrogen Oxide—NOX 
                 10.97 
                 10.97 
               
               
                   
                 Carbon Monoxide—CO 
                 0.95 
                 1.07 
               
               
                   
                 Hydrocarbons—HC 
                 21.9 
                 2.39 
               
               
                   
                 Carbon Dioxide—CO2 
                 2.1 
                 2 
               
               
                   
                 Oxygen—O2 
                 1.41 
                 1.43 
               
               
                   
                   
               
             
          
         
       
     
     G1 Fuel Usage Results Per Unit of Brake Horsepower 
     
         
         
           
             Low Load Fuel Usage: 10% Less than Factory Engine (12.07−10.86=1.21/12.07) 
           
         
       
    
     
       
         
               
               
               
               
             
           
               
                   
                   
               
               
                   
                   
                 FACTORY 
                   
               
               
                   
                 HIGH LOAD 
                 ENGINE 
                 G1 ENGINE 
               
               
                   
                   
               
             
             
               
                   
                 Load % 
                 80% 
                 80% 
               
               
                   
                 RPM 
                 3550 
                 3550 
               
               
                   
                 Run Time 
                 1:30 minutes 
                 1:30 minutes 
               
               
                   
                 lb-ft Torque 
                 10 
                 14 
               
               
                   
                 Brake Horsepower 
                 6.7 
                 9.4 
               
               
                   
                 Fuel Usage - Milliliters 
                 13.19 
                 8.65 
               
               
                   
                 Nitrogen Oxide—NOX 
                 5.97 
                 4.57 
               
               
                   
                 Carbon Monoxide—CO 
                 0.58 
                 0.44 
               
               
                   
                 Hydrocarbons—HC 
                 11.04 
                 1.07 
               
               
                   
                 Carbon Dioxide—CO2 
                 1.29 
                 0.8 
               
               
                   
                 Oxygen—O2 
                 1.34 
                 0.67 
               
               
                   
                   
               
             
          
         
       
     
     G1 Fuel Usage Results Per Unit of Brake Horsepower 
     
         
         
           
             High Load Fuel Usage: 34.4% Less than Factory Engine (13.19−8.65=4.54/13.19) 
           
         
       
    
     G1 High Load Emission Results Per Unit of Brake Horsepower 
     
         
         
           
             NOX: 23.4% Less than Factory Engine HC: 90.3% Less than Factory Engine 
             CO: 24.1% Less than Factory Engine CO2: 37.9% Less than Factory Engine 
           
         
       
    
     Two GX 390 Honda 13 hp engines were used for testing and comparisons (i.e., a “stock” engine versus one modified in accordance with the instant invention). Both engine specifications were as follows:
         Four stroke valve single cylinder   3.5×2.5 bore &amp; stroke   4.412 rod length   Forced air cooling systems   Gravity feed fuel systems   87 octane gasoline   23.7 cu/in displacement   Transistorized magnet ignition systems       

     The muffler was removed on both engines to confine exhaust emissions for analysis purposes. The engine with the stock head is named the “Factory” engine on the above chart. The engine with our proprietary head is named the “G1” on the above chart. 
     All tests were conducted on the same day in a controlled and isolated environment. Fuel and emission measurements were made using the following equipment:
         Land &amp; Sea Water Brake Dyno, the Dyno-Max 2000 Model   Dyno-Max 2000 Data Analysis Software and Multimedia PC Demonstration, 9.38 SPI Version   UEI AGA 5000 Emissions Analyzer   ASTME rated ⅜ inch Bellwether 100 cc Tube       

     The primary objective of house testing was to determine the fuel usage of the modified engine. We kept run time, load and rpm constant. To compare and measure the efficiency, input was divided by output. In our particular case, fuel usage was our input variable and our output variable was the pound-foot of torque produced. Fuel usage and all emissions results of both engines were calculated based on a unit of brake horsepower (torque×rpm/5252). 
     The low load fuel usage per unit of brake horsepower for the G1 engine was 10% less than the Factory engine. The high load fuel usage per unit of brake horsepower for the G1 engine above. It was determined that fuel consumption of the modified engine G1 was 34.4% less than the Factory engine. The high load emissions per unit of brake horsepower for the G1 engine resulted in 23.4% less nitrogen oxide (NOX), 
     24.1% less carbon monoxide (CO), 90.3% less hydrocarbons (HC) and 37.9% less carbon dioxide (CO2) compared to the Factory engine. 
     From the foregoing, it will be seen that this invention is one well adapted to obtain all the ends and objects herein set forth, together with other advantages which are inherent to the structure. 
     It will be understood that certain features and subcombinations are of utility and may be employed without reference to other features and subcombinations. 
     As many possible embodiments may be made of the invention without departing from the scope thereof, it is to be understood that all matter herein set forth or shown in the accompanying drawings is to be interpreted as illustrative and not in a limiting sense.

Technology Classification (CPC): 5