Rotating barrel type internal combustion engine

An internal combustion barrel engine having rotating cylinders and pistons which together form combustion spaces. The combustion spaces are maintained at a substantially constant volume while a compressed air-fuel mixture is combusted therein.

BACKGROUND

The present invention relates to engines of all sorts. More particularly, the present invention relates to an internal combustion engine of a barrel-type configuration in which the cylinder axes are arranged around a central longitudinal axis of the engine, and even more particularly to a barrel-type engine having a rotating cylinder bank.

Internal combustion engines have been around for a long time. The basic components of the engine are well known in the art and include the engine block, cylinder head, cylinders, pistons, valves, crankshaft and camshaft. The cylinder heads, cylinders and tops of the pistons typically form combustion chambers into which fuel and air are introduced so that combustion takes place. Useful work is generated from the hot, gaseous products of combustion acting directly on the top or crown surface of the piston. Generally, reciprocating linear motion of the pistons within the cylinders is transferred to rotary motion of a crankshaft via connecting rods. One common internal combustion engine is known as an Otto-type internal combustion engine and employs a four-stroke cycle in which power is derived from the combustion process over four separate pistons movements (strokes): intake stroke, compression stroke, expansion (power) stroke, and exhaust stroke. In traditional Otto-type automotive engine applications, the cylinders are typically stationary and are typically arranged in one of three ways: (1) a single row (in line) with the centerlines of the cylinders commonly vertically oriented; (2) a double row with the centerlines of opposite cylinders converging in a V (V-engine); or (3) two horizontal, opposed rows (opposed or pancake engine). Two additional Otto-type cylinder configurations were also experimented with, primarily between 1900 and 1950, and include (1) a radial configuration where the cylinder axes are arranged like spokes of a wheel with the lower rod ends mounted on a common crank shaft journal, and (2) a barrel configuration with cylinder axes arranged parallel around the central longitudinal axis of the engine. Barrel configurations generally include a stationary cylinder bank and the power is transferred to the crankshaft in one of three ways (1) with the lower ends of the connecting rods connected to a gear arrangement, (2) with the lower ends of the crankshaft connected to a wobble plate, and (3) with the lower ends of the rods pushing a cam surface.

A subclass of barrel engines are those with a rotating cylinder bank and such engines generally come in one of three configurations: (1) a two or four-cycle arrangement in which the rotating cylinder bank drives an angled thrust plate from which power is taken off as shown by way of example in U.S. Pat. Nos. 980,491; 1,345,808; 2,382,280 and 4,779,579; (2) a two-cycle arrangement in which a pair of rotating cylinder banks share a common cylinder head unit and in which the outer rod ends each drive an angled thrust plate as shown by way of example in U.S. Pat. Nos. 968,969; 1,255,664 and 1,779,032; and (3) a two-cycle arrangement in which a pair of rotating cylinder banks share a common piston and in which a pair cylinder head units are provided at each end thereof as shown by way of example in U.S. Pat. Nos. 3,830,208 and 5,103,778. It is believed, both radial and barrel engines, in particular, fell out of favor after World War II.

Beginning in the early part of the twentieth century, the conventional Otto-type reciprocating engine began to assume dominance as the most practical approach, even though it was recognized that the thermodynamic efficiency of the engine was such that about two-thirds of the energy developed through the combustion of the fuel was wasted. That is, roughly ⅓ of the fuel energy is delivered to the crankshaft as useful work, ⅓ is lost in waste heat through the cylinder walls, heads and pistons, and ⅓ is lost out of the exhaust.

The Wankel engine, which is also known as a rotary engine, is denoted as such because it utilizes a single triangular rotating piston which forms combustion chambers as it rotates within a stationary figure eight-shaped “cylinder”. The Wankel engine does not employ connecting rods as the rotating piston is linked directly to the crankshaft. The Wankel engine is also a four-stroke cycle engine, and while it has several advantages over the Otto-type engine, it produces higher emissions, has a shorter lifespan, and lacks torque at low speeds, which leads to greater fuel consumption.

Applicant's U.S. Patent Application Publication No. 2003/0131807 provides an improved barrel configuration with a rotating cylinder bank and angled thrust plate. However, it is always desirable to make improvements such as but not limited to improvements in thermodynamic efficiency, emissions, manufacturability, and/or power or torque of the engine.

SUMMARY

The Summary and Abstract are provided to introduce a selection of concepts in a simplified form that are further described below in the Detailed Description. The Summary and Abstract are not intended to identify key features or essential features of the claimed subject matter, nor are they intended to be used as an aid in determining the scope of the claimed subject matter. In addition, the claimed subject matter is not limited to implementations that solve any or all disadvantages noted in the Background.

An aspect of the present invention is an internal combustion barrel engine having rotating cylinders and pistons which together form combustion spaces. The combustion spaces are maintained at a substantially constant volume while a compressed air-fuel mixture is combusted therein. Using various design orientations, relationships, positions, tilts and/or offsets of the rotating cylinders and thrust plate to which the pistons are connected, a dwell can be obtained where the piston remains substantially stationary with respect to the corresponding cylinder when transitioning from a compression stroke to a power stroke and/or control the speed of the piston during various portions of the cycle.

In one embodiment, an engine block assembly includes a stationary housing, a cylinder bank rotatably mounted to the housing about a central longitudinal axis, the cylinder bank having a plurality of cylinders therein radially distanced from the central longitudinal axis, each cylinder having associated therewith a cylinder wall formed about a major cylinder axis, a plurality of pistons wherein one piston is provided in each cylinder to form a combustion chamber therein, wherein each piston sequentially moves from a down most position within the cylinder to an up most position within the cylinder during a first portion of rotation of the cylinder bank, wherein each piston sequentially dwells about the up most position for substantially all of an air-fuel mixture to be combusted within the combustion chamber, and wherein each piston then sequentially moves from about the up most position to the down most position during a second portion of rotation of the cylinder, a plurality of connecting rods each having a proximal end attached to a respective piston, and a remote end distant from the respective piston, a thrust plate operatively connected to the remote ends of the connecting rods, the thrust plate being rotatably mounted to the stationary housing about a thrust plate axis and in a thrust plane defined by the remote ends of the connecting rods, a synchronizing member operatively connecting to the cylinder bank and the thrust plate so that the cylinder bank and thrust plate rotate at the same speed. The piston dwell motion is created by adjusting one or more of the following design parameters: (1) the angle of the thrust plane with respect to a plane that is perpendicular to the central longitudinal axis, (2) the angular rotational offset of the thrust plate about an axis which is parallel to the central longitudinal axis and which intersects the thrust plate axis, (3) the angular rotational offset of the thrust plate about the thrust plate axis with respect to a reference point in the thrust plane, (4) the lateral offset of the thrust plate axis from the central longitudinal axis, and (5) the tilt of the major cylinder axes with respect to the central longitudinal axis.

These and other aspects will be described further below.

DETAILED DESCRIPTION

In the description below various exemplary embodiments of engines will be described. It should be understood that aspects of the exemplary embodiments are not limited to the embodiment in which such aspects are described, or in other words, such aspects can be included on any other exemplary embodiment herein described or other embodiments beyond those described, if desired. Where relevant in the description references will be made to the various embodiments when describing similar or alternative aspects, components or mechanisms.

FIGS. 1 and 2illustrate an exemplary rotating four-cycle barrel type internal combustion engine10having aspects of the present invention. Other embodiments are provided below. In the exemplary embodiment, engine10includes a stationary housing assembly11, rotating cylinder bank assembly12for power generation, a power take-off assembly14for generating torque, a fuel delivery system16(FIG. 5) for regulating the fuel intake to the engine10, a scavenging system18to minimize engine emissions, an air delivery system20for charging the fuel, cooling the cylinder bank assembly12and scavenging, an ignition system22for igniting the fuel, and a liquid cooling system as represented by passageway24(FIG. 2) for cooling the cylinder bank assembly12. It should be understood that aspects of the present invention are not limited to an engine having all parts to operate. For instance, aspects of the present invention can be included in an engine block assembly having, for example, cylinders and pistons with or without a power take-off assembly or other subsystems such as a fuel delivery system, ignition system, cooling system, air delivery system, etc. As appreciated by those skilled in the art these and other subsystems can take any number of forms in order to provide an operable engine.

In the exemplary embodiment, a four-stroke cycle operation is provided in the course of two complete revolutions of the engine10as follows: an intake stroke ranging from about 0□ to about 180□ of the first revolution of the engine10, a compression stroke ranging from about 180° to about 360° of the first revolution, a power stroke ranging from about 360° to about 540° of the second revolution, and an exhaust stroke ranging from about 540° to about 720° of the second revolution. It should be noted that the aforementioned and following degree ranges are for purposes of understanding only. The degree ranges may be adjusted to affect the power, speed, torque, fuel economy and/or emission quality for each application of the engine10.

The stationary housing assembly11houses and secures the engine in a relative stationary position such as, but not limited to, for pumps or generators, or in a vehicle (not shown, but without limitation including any vehicle operable on/in land, water and/or air). The housing assembly includes a combustion exhaust manifold30, a cylinder head cooling exhaust manifold32, a cylinder cooling exhaust manifold34, and a pair of scavenging exhaust manifolds36and37(FIG. 6). A seal38(FIG. 1) within the combustion exhaust manifold30prevents exhaust fumes from leaking out of the manifold34. A back pressure passageway40provides air at a higher pressure than the exhaust gases to ensure that exhaust gases do not leak past the seal38. The manifolds30,32, and34can have longitudinal cooling fins extending from an exterior thereof to provide both improved heat transfer and improved structural support. The combustion exhaust manifold30is exposed from about 185° to about 350° to coincide with the exhaust stroke of the engine10. The cylinder head exhaust manifold32and cylinder cooling manifold34can be exposed during the entire 360°. revolution of the engine, and the heated air stream generated may be used for other purposes such as to heat a passenger compartment of the vehicle. The combustion exhaust manifold30, the cylinder head cooling exhaust manifold32, and the cylinder cooling exhaust manifold34may be spiraled to more efficiently remove the gases from the engine10. Referring also toFIG. 6, the scavenging system18includes a stationary pre-exhaust scavenging manifold36positioned near bottom dead center of the engine for directing unburned fuel scavenged from the cylinder bank assembly12back into the fuel delivery system to improve emissions, and a post exhaust scavenging manifold37positioned near top dead center of the engine for directing all residual burned fuel scavenged from the cylinder bank assembly12back into the fuel delivery system16to improve emissions, as will be further explained below.

The cylinder bank assembly12is rotatably mounted to the stationary housing11about a central longitudinal axis42and for example using suitable bearings such as bearings44and45. The cylinder bank assembly12includes a plurality of cylinders46each having an upper end47, a lower end48and a cylinder wall49, a cylinder head assembly50mounted to the upper end47of the cylinders46for rotation therewith, a cylinder carriage52mounted to the lower end48of the cylinders46for rotation therewith and having a synchronizing gear53thereon for transferring torque to the power take-off assembly14and a starter gear55on a peripheral surface thereof, a plurality of pistons54each of which is moveable within a respective one of the plurality of cylinders46between an up position and a down position as the cylinder bank assembly12rotates, a plurality of connecting rods56each of which has an inner end57connected to the underside of a respective one of the plurality of pistons54and an outer end58operatively connected to the power take off assembly14via retainers59so that the outer end58of the rod56freely rotates and pivots as necessary as the cylinder bank assembly12rotates. The pistons54can each have a partial skirt65extending from an underside thereof and providing an improved wear surface against the cylinder wall49while at the same time minimizing piston weight. The cylinder walls49can have a corresponding partial skirt67for supporting the pistons skirt65and at the same time minimizing weight of the rotating mass. Centripetal force of the rotating cylinder bank assembly12should keep the piston skirts65oriented towards the outside of the cylinders46where the wear is greatest. Should the pistons rotate within the cylinder as the cylinder bank rotates than it would be desirable to use a fully skirted piston rather than the partial skirt65. A starter motor61(FIG. 2) operatively connected to the stationary housing11includes a gear63which meshes with the starter gear55on the cylinder carriage52for initiating rotation of the cylinder bank assembly12.

The cylinder head assembly50includes a head unit60having an intake port62and an exhaust port64positioned adjacent to each of the plurality of cylinders46, a valve assembly66for opening and closing the intake port62and the exhaust port64to the cylinders in a timed sequence, and a cam assembly68for controlling the valve assembly66. The head unit60is shown dough-nut shaped having an inner surface70, an outer surface71, an upper surface72and a lower surface74. With respect to each cylinder, the lower surface74of the head unit60includes a domed shaped valve seat75separating the intake and exhaust ports62and64from the cylinders and a wall76separating the intake port62from the exhaust port64from each other. The valve assembly66controls the opening and closing of the intake port62and the exhaust port64with respect to the cylinders46by sealing against the valve seat75. The combustion exhaust manifold30controls access to the exhaust port62while the fuel delivery system16controls access to the intake port64.

The valve assembly66includes a valve80, a valve lifter81, a valve return spring82, a tracking roller83, and a retainer84. Each valve80is disposed in the head unit60for sealing a respective cylinder46from the intake port62and the exhaust port64thereof and is built to withstand the full pressure of the expanding gasses within the combustion chambers. The valves80can be poppet valves as are used in standard contemporary gasoline engines. This single valve configuration can be advantageous over separate intake and exhaust valves because it achieves greater volumetric efficiency, simplifies the cam geometry, enables less energy to be spent depressing the valve only once during each four cycle operation, and reduces the need for rapid acceleration of the valve stroke as is necessary in a two valve configuration. Nonetheless, it is intended that the spirit and scope of this invention extend to an embodiment with separate intake and exhaust valves and actuation thereof. Each valve80includes a stem86operatively connected to a proximal end of the valve lifter81via the valve return spring82which biases the valve80in a closed position. The retainer84keeps the tracking roller83engaged to a distant end of the valve lifter81. The tracking roller83is positioned at the upper surface72of the head unit and engages the cam assembly68for moving the valve80up and down and thereby controlling the closing and opening of the intake port62and exhaust port64of respective cylinders46.

The cam assembly68includes a cam plate90adjacent the upper surface72of the head unit60and having a plurality of cam surfaces92protruding therefrom, or other mechanical actuator which controls the valves80, so as to open each valve80commencing at the exhaust stroke (about 540° to about 720°) and remain open through the intake stroke (about 0° to about 180°) and so as to close each valve80commencing at the compression stroke (about 180° to about 360°) and remaining closed throughout the power stroke (about 360° to about 540°). It can be advantageous to use an odd number of pistons54and corresponding cylinders46so that every other piston54continuously fires while the cylinder bank assembly12is rotating in normal four-cycle operation. The cam plate90has an internal gear94that engages an external gear96on the rotating cylinder bank assembly12at one position as shown inFIGS. 1 and 2. The cam plate90is rotatably mounted to the stationary housing11about a cam axis98such as by bearings44and45. The cam axis98is essentially parallel to the central longitudinal axis42and radially offset outwardly from it in the direction corresponding to bottom dead center of each piston54in its corresponding cylinder46. This offset can be determined by the difference in the radius of the gears94and96on the spinning cam plate90and the rotating cylinder bank assembly12, respectively. The cam plate90spins at an exact synchronous ratio to the cylinder bank assembly12so that the cam surfaces92are timed to actuate the valves80according to the particular timing sequence of the engine10. Cam surfaces92can be similar to cam surfaces described in U.S. Patent Application 20030131807 entitled “Rotating Positive Displacement Engine”, and published Jul. 17, 2003, incorporated herein by reference in its entirety.

In the illustrated example of a seven-cylinder engine, it is preferred that the cam plate90rotate slower than the cylinder bank assembly12so that the cam plate90advances seven rotations for every eight rotations of the cylinder bank assembly. The seven-to-eight gear ratio causes each valve80to be opened only for the desired fuel exhaust and intake cycles of the engine10, and to remain closed for the compression and power cycles of the engine10. In this arrangement there is provided four protruding cam surfaces92on the cam plate90. The profile of the cam surfaces92as well as the area between the cam surfaces92are shaped so that with the seven-to-eight gear ratio of the cam plate90to cylinder bank assembly12and with the axial offset therefrom, the cam surfaces92uniformly contact and stay in uniform contact with all of the tracking rollers83as the cylinder bank assembly12rotates. Depression of the tracking roller83by the cam surfaces92thereby depresses the respective valve lifer81and corresponding valve80as the engine rotates, so that each valve80is depressed only one time for a period of approximately 360° in every two rotations (720°) of the cylinder bank. The valve return spring82returns the valve80to the closed position after the cam surface92moves past the tracking roller83. For other design embodiments involving a different odd number of cylinders46(for example 1, 3, 5, 9, 11, etc.) and a different number of valves80per cylinder46(for example 1, 2, 3, 4, etc.) there will be a different timing ratio and a different number of cam surfaces92on the cam plate90. For example,FIGS. 17 and 18illustrate a five cylinder engine10′ having two valves per cylinder (intake valve80A and exhaust valve SOB) and two cam plates (intake valve cam plate90A and exhaust valve cam plate90B) offset with respect to each other to actuate an intake valve80A and an exhaust valve80B, respectively. In such an arrangement, each of the two cam plates90A,90B would spin slower than the cylinder bank12′ at a ratio of 5/6 its speed and there would be six cam surfaces on each cam plate90A,90B so that each respective intake valve80A and exhaust valve BOB is actuated along each of the respective cam surfaces during the course of six revolutions of the cylinder bank12′. In this embodiment, each valve80A,80B is operated through a roller83A,83B that contacts the corresponding cam plate90A,90B. Each roller83A,83B is supported on a push rod83C that in turn actuates a rocker83D that operates the corresponding valve BOA, SOB. In this case the contact speed of each roller83A,83B to the corresponding cam plate90A,90B is ⅙ the engine speed. Referring to the embodiments ofFIGS. 1-2and17-18, while it is possible to spin the cam plate90,90A,90B faster than the cylinder bank assembly12,12′ and achieve proper synchronization, it is advantageous to spin the cam plate90,90A,90B at a slower speed to minimize impact of the tracking rollers83,83A,83B against the corresponding cam surfaces. It should also be noted with regard toFIGS. 17 and 18that the cam surfaces may be located on the lateral edge of the generally flat star-shaped cam plate90A,90B, and as such the flat cam plates90A,90B are more easily machined than the cam plate90shown inFIG. 1. The cam plates90A,90B (which can be formed from an integral unitary body) include gear teeth94′ that mate with a drive gear96′ that rotates with the cylinder bank12′.

As described above, conventional rollers83A,83B moving along the lateral or perimeter edge cam surface actuate conventional rockers83D, lifters and springs to open and close the corresponding valves80A,80B. In should be noted that the star-shaped cam plates90A,90B shown inFIGS. 17 and 18appear as flat surfaces to the rollers83A,83B and that identical cam lobes (not shown) would be positioned on each lateral edge of the six-sided star-shaped intake cam plate90A, and another set of identical cam lobes (not shown) would be positioned on each lateral edge of the exhaust cam plate908. The cam lobes are not shown because their position is determined by the desired valve timing. Referring back to the exemplary embodiment ofFIGS. 1-2, the air delivery system20includes a primary air compressor102and a secondary air compressor104and is used to cool the engine10and to compress or supercharge the fuel-air mix for increased combustion. The primary air compressor102is rotatably mounted via bearings on a first drive shaft106which is substantially aligned with the central longitudinal axis42and the secondary air compressor104is rotatably mounted on a second drive shaft108which is concentric within the first drive shaft106. An inner end109of the second drive shaft108is rotatably mounted to the cylinder carriage52for support. The primary and secondary air compressors102and104spin independently at different speeds with respect to each other and at a substantially greater rate than the cylinder bank assembly12. The primary and secondary air compressors102and104are driven by any one of a variety of methods including a gear train (not shown) directly linked to the rotating cylinder bank assembly12. The air compressors102and104can also be driven by variable speed electric motors110and111(FIG. 2), respectively, which transfer power either directly or through a power train. The speed of the electric motors110and111is variable and governed by a control unit112via a connection line so as to control the pressure and volume of air provided to the engine10in proportion to the needs of varying operating engine conditions such as load, rpm, temperature, acceleration, etc.

The engine conditions are monitored through the use of dedicated real time sensors (not shown), which are well known in the art, for measuring conditions such as rpm, load, throttle position, cylinder temperature, head temperature, air velocity, exhaust composition, and manual override, etc. However, it may be desirable to use optical or radio frequency transmission for sensors which are placed on-board the rotating cylinder bank. One of the uses for the compressed air can be to cool the cylinders46and the head unit60. As shown inFIG. 3, in regard to the cylinders46, the cylinder walls49have a plurality of cooling fins114extending out therefrom in a respective plurality of planes each of which are substantially perpendicular to the central longitudinal axis42and are cut to form a lateral wedge-shaped cooling fin arrangement116which communicates with the lateral wedge-shaped cooling fin arrangements116of adjacent cylinders46to provide maximum heat transfer surface area. The cylinder carriage52acts as a baffle directing pressurized air flowing down the center of the engine10out across the cylinder cooling fins114. Referring toFIGS. 2 and 4, in regard to the head unit60a plurality of cooling slots118are located on the inner surface70thereof and a plurality of cooling fins119arranged on the outer surface71thereof. Ambient air flows axially and radiates downwards from the air intake port in the primary air compressor102towards the circumference of a stationary compressor shroud120by action of compressor impellers122and thereby becomes pressurized for entering the rotating head unit60where it is then directed through the plurality of cooling slots118and across the cooling fins119for cooling the head unit60. A portion of the pressurized air passes down through the center of the cylinder head60and into the fuel delivery assembly16where it is further pressurized by the secondary air compressor104. A first portion of this further compressed air then passes through an opening117(seeFIG. 5) below the secondary air compressor104and into the lateral wedge-shaped cooling fin arrangements116for cooling the cylinders46as described above. A second portion of this further compressed air is directed into the fuel delivery system16to create an air-fuel mixture and then into the plurality of cylinders46for combustion. Alternatively, this second portion of further compressed air may be delivered into the plurality of cylinders46without the fuel so as to provide compression resistance within the cylinders to slow the engine speed. A third portion of this further compressed air is used in the scavenging system18. A fourth portion of this further compressed air is used to back pressure the combustion exhaust manifold30via back pressure passageway40.

Referring toFIGS. 2-4, the liquid cooling system24provides added cooling of the cylinder bank assembly12by way of least one closed-loop passageway124self contained within the cylinder bank assembly12, wherein each passageway124has a hot area125and a cooler area126A; and a heat expansive liquid contained within the closed-loop passageway124for transferring heat from the hot area125to the cooler area126A as the cylinder bank assembly12rotates. More specifically, the head unit60further includes a plurality of closed loop passageways124therein, each passageway124having a hot area125adjacent the valve80and a cooler area126A distant to the valve80. The heat expansive liquid within the passageway124transfers heat from the hot area125to the cooler area126A along a toroidal path as the head unit60rotates. The liquid flow is caused via a centripetal force acting on the heat expansive liquid as it becomes less dense moving to the hotter area125. The centripetal force caused by the rotating head unit60causes this more dense material to move outward from the heat source thereby effectively transferring the heat. The cooler area126A of the passageway124slopes towards a perimeter of the head unit60and the hotter area125of the passageway124slopes towards an interior of the head unit60to create a toroidal flow within the passageways124. The cooling fins119of the head unit60extend radially out from the side walls of the cooler area126A of the passageways124, The cooling slots118(FIG. 4) of the head unit60are positioned between each passageway124so that cooling air passes across the cooling fins119. Holes128(FIG. 4) in each of the cooling fins119permits air to flow between all of the cooling fins119for added air circulation. Cooling air from the primary air compressor102passes through cooling slots118, moves across the side walls of the passageways124and then across and between the cooling fins119to provide cooling of the head unit60.

In addition or in the alternative to passageways124described above, each of the plurality of cylinders46can also have at least one closed-loop passageway134self contained adjacent the cylinders walls49of each of the plurality of cylinders46. Each of the closed-loop passageways134is adjacent hot area125and has a cooler area126B; and a heat expansive liquid contained within the closed-loop passageway134for transferring heat from the hot area125to the cooler area126B as the cylinder bank assembly12rotates. More specifically, each cylinder46further includes an upper chamber130adjacent to the upper end47of the cylinder46acting as the hot area125, a lower chamber132adjacent to the lower end48of the cylinder46acting as the cooler area126B, and a plurality of tubular passageways134connecting the upper chamber130to the lower chamber132so that the heat expansive liquid flows in a toroidal manner from the cooler areas126B to the hotter areas125and vice versa. The tubular passageways134are angled so that the heat expansive liquid within the passageway134transfers heat from the hot area125near the valve80to the cooler area136at the distant radius of the lower end48of the cylinder46. An oblique angle of the tubular passageways134allows the centripetal force to move the colder more dense liquid at the lower end48of the cylinders46upwards towards the periphery of the cylinder bank and the valve80where it then becomes hotter and less dense so that it then moves inwards towards the center of the cylinder bank causing a toroidal flow effectively transferring heat and cooling the cylinders46. The cylinder cooling fins114extend across an exterior surface of the tubular passageways134so that cooling air from the primary and secondary compressors102and104passes over the exterior surface of the passageways and across the cooling fins114to cool the cylinders46. Cylinder cooling fins114also extend out from the cylinder wall49within the upper chamber130and within the lower chamber132to aid in heat transfer. It is desirable to connect the closed-loop passageways124between the head unit60and the cylinders46to each other to further aid in cooling. The closed-loop liquid cooling system24described herein is desirable because it does not require any external energy source other than the rotating motion of the cylinder bank assembly12. In addition, because the system24is self-contained within the rotating cylinder bank assembly12sliding seals and additional bearings are not needed as would be the case if the cooling liquid is pumped in from an external radiator. Nonetheless, it may be desirable or required to pump the heat expansive liquid to an external radiator to increase the volume of the fluid flow and provide adequate heat transfer.

Referring toFIGS. 2 and 5, the fuel delivery system16includes a fuel supply unit136, one or more fuel lines138which extend from the fuel supply unit136and pass through a portion of the stationary housing11, a series of liquid fuel injectors140connected thereto for mixing and admitting atomized liquid fuel to the pressurized air, and a throttle142for controlling the amount of fuel/air mixture that is admitted to the cylinders46. The control unit112regulates the amount of fuel admitted to the fuel injectors140as well as the operation of the throttle142and the speed of the air compressors102and104. The fuel injectors140are of the common rail type and are well known in the art. The throttle142, on the other hand, includes a stationary throttle support144fixedly mounted to the stationary housing11, an actuator146having a first arc-shaped door147and an actuator gear148thereon, a second arc-shaped door150having an actuator gear151thereon, a cylinder head interface barrier152rigidly attached to the stationary throttle support144for providing the interface between the first and second doors147and150and the intake ports62of the cylinders46and for providing a fuel administration opening154therethrough, a synchronizing pinion gear156rotatably mounted to the stationary throttle support144for simultaneously moving the first and second throttle doors147and150either away from each other to increase flow through the fuel administration opening154or towards each other to decrease flow through the fuel administration opening154, an actuator pinion gear158rotatably mounted to the stationary throttle support144for engaging the actuator gear151, and a control unit112which controls the actuator pinion gear151through a rod159via line160. The stationary throttle support144includes a head portion161which provides the offset for the cam axis98, a neck portion162which has a plurality of cooling slots164thereon for directing pressurized air from the primary air compressor102to the head unit60, and a base portion166into which the stationary fuel injectors140are fixedly mounted so as to admit atomized fuel into a stream of air moving into each of the plurality of the cylinders46during the intake stroke thereof, for example, in sequence as cylinders pass by on their respective intake stroke. The actuator146is constructed in two pieces so as to be rotatably mounted around the neck portion162of the stationary throttle support144about the central longitudinal axis42. The actuator146includes a neck portion168having a plurality of cooling slots169thereon for directing pressurized air passing through the cooling slots164of the neck portion162of the stationary throttle support144to the head unit60for cooling thereof, and a base portion170having an opening172allowing the actuator146to rotate around the stationary fuel injectors140. The first door147extends outward from an underside of the base portion170in an arc shape and circumferentially moves with respect to the arc shaped second door150through the pinion gear156so as to open and close an arc shaped opening along the entire circumferential arc forming the intake stroke. It is important to note that the arc shaped opening exposed by circumferential movement of the throttle doors147and150can be increased or decreased both radially and along a circumferential arc defined by the intake cycle, thereby providing maximum control in delivering air and air-fuel mixture to the cylinders46. In the case of an engine10having seven or more cylinder, the throttle simultaneously delivers air and air-fuel mixture to at least two open cylinders46during the entire intake cycle.

FIGS. 19-21illustrate an alternative fuel delivery system16′ in which there is a stationary semi circular manifold173A mounted to the stationary housing represented by support shaft175for example with spokes not shown, and a rotating semi-circular manifold173B mounted to the rotating cylinder bank12′ and which nests with the stationary semi circular manifold173A. The stationary manifold173A is only exposed on the intake side of the engine and is closed off on the exhaust side. The rotating manifold173B includes separate runners or passageways173C leading to each of the intake valves80A of the cylinders. InFIG. 20, common rail fuel injectors140are positioned in the stationary semicircular manifold173A and controlled as described above so that a controlled amount of fuel is delivered to the cylinders. Seals177are used between the stationary manifold173A and the rotating manifold173B to prevent escape of the fuel air mixture and it may be desirable to use a small blower to back pressure the seals.

Referring toFIG. 2, the ignition system22includes a plurality of spark plugs174arranged singular or in pairs on both sides of the valve80associated with each cylinder46, a pair of spark plug contact strips176connected to each of the spark plugs174within each cylinder46, a spark plug commutator178mounted to the stationary housing assembly11so as to operate in contact with the spark plug contact strips176as the head unit60rotates, and the control unit112for providing the desired ignition timing and sequence. The fuel delivery system22admits a fuel and air mixture in a timed sequence into each cylinder46via its intake port62as the piston54therein moves from an up position to a down position as the cylinder bank assembly12rotates. The fuel/air mixture is then compressed within the cylinder46as the piston54therein moves from the down position to the up position as the cylinder bank assembly12rotates, and then the control unit112ignites the fuel/air mixtures in timed sequence as the spark plugs in each cylinder operatively engages the spark plug commutator at some point before top dead center so that the flame kernel can fully develop when the piston has maximum mechanical advantage. The spark plug contact strips176have independent metal contact strips connected to each of the spark plugs174for independently and simultaneously firing both spark plugs174within each cylinder46. The relatively slow formation of the initial flame kernel and the subsequent burn produces a peak cylinder pressure after top dead center. The explosion drives the respective piston54from the up position to the down position and causes the power take off assembly to rotate thereby creating torque. The combusted gases within the cylinder46are exhausted through the exhaust port64thereof and into the combustion exhaust manifold30as the piston moves from the down position to the up position. In order to achieve the four-cycle operation, it is preferred that there is an odd number (1, 3, 5, 7, 9, etc.) of combustion chambers so that as the cylinder bank assembly12rotates, each cylinder46goes through the four-cycle operation in a simple timed sequence wherein every other cylinder46is acted upon. More specifically, on one side of the engine10adjacent cylinders46alternate between the intake and power cycles, wherein the control unit112times the spark plugs174so as to fire in every other cylinder46as the cylinder bank assembly12rotates, and wherein the fuel control assembly14admits a fuel and air mixture to every other cylinder46as the cylinder bank assembly12rotates. On the other side of the engine10, the adjacent cylinders alternate between the compression and exhaust cycles. In the seven cylinder engine, this alternate firing/fueling and, conversely, compression/exhaust provides continuous operation and accomplishes the four-cycle operation for all of the cylinders46in the course of two full rotations of the cylinder bank assembly12in the following sequence: Cylinder #1, #3, #5, #7, #2, #4, #6, #1, etc.

Referring toFIG. 6, the scavenging system18is provided to minimize emissions and maximize efficiency of the engine. The scavenging system18includes a pre-exhaust scavenging system180which scavenges any residual fuel which gets trapped in the cylinder intake ports62and exhaust ports64after the valve80closes so as not to leak unburned fuel into the combustion exhaust manifold30, and a post exhaust scavenging system182to scavenge any residual combustion exhaust out of the cylinders46before commencing the intake stroke. The pre-exhaust scavenging system180operates on each and every intake and exhaust port62and64at approximately bottom dead center when the cam assembly68is transitioning the valves80from closed to open (to commence the exhaust stroke) or from open to closed (to commence the compression stroke). At bottom dead center all valves80are closed which is just before a leading edge of one cam surface92opens a valve80whose cylinder46is about to start exhaust and just after a trailing edge of an adjacent cam surface92falls off causing the adjacent valve80to close after the intake stroke. Air from the secondary air compressor104is bled off through a pre-exhaust scavenging opening184in the stationary throttle support144, through a pre-exhaust scavenging opening185in the second throttle door150, through a pre-exhaust scavenging opening186in the cylinder head interface barrier152, through the intake and exhaust ports62and64for scavenging, out into the stationary pre-exhaust scavenging manifold36which directs the scavenging gases up, around and down through the stationary throttle support144so as to recycle the scavenged gases into the secondary air compressor104adjacent to the intake ports62for charging the cylinders46during the intake stroke. The post exhaust scavenging system182also operates with respect to each and every cylinder46except that some valves80are open and some are closed depending on whether the cylinder46is ready to transition from the compression stroke or the exhaust stroke. The post exhaust scavenging system182is portioned adjacent top dead center when the valve80of cylinders46in the exhaust stroke is still open and when the exhaust port64is out of communication with the combustion exhaust manifold30, before the intake ports62are exposed for charging of the cylinders46. With respect to closed valve cylinders46, air from the secondary air compressor104is bled off through a post exhaust scavenging opening188in the stationary throttle support144, through a post exhaust scavenging opening189in the first throttle door147, through a post exhaust scavenging opening190in the cylinder head interface barrier152, through the cylinder intake and exhaust ports62and64for scavenging, out into a stationary post-exhaust scavenging manifold37which directs the scavenging gases up, around and down through the stationary throttle support144so as to recycle the scavenged gases with the pre-exhaust scavenged gases and into the secondary air compressor104adjacent to the intake ports for charging the cylinders during the intake stroke. With respect to open valve cylinders46, air from the secondary air compressor104passes through post exhaust scavenging openings188,189and190, through the cylinder intake port62, into the cylinders46where it swirls down and then out through the exhaust port64scavenging any residual combustion exhaust gases into a stationary post-exhaust scavenging manifold37as indicated above.

Referring toFIGS. 5 and 6, a water injector192may be provided for added cooling of the valve80on demand. The water injector192is mounted into the stationary throttle support144and positioned adjacent to the post exhaust scavenging opening188for squirting atomized water directly onto the valve80for added cooling, if needed, and for adding to the density of the scavenged gases which enter the stationary post-exhaust scavenging manifold37. The water injector192is connected to the control unit112via line193so as be activated as engine conditions demand.

Referring toFIGS. 1 and 2, in its simplest form the power take off assembly14includes a load bearing thrust plate200having a synchronizing gear202thereon, a stationary thrust housing plate204, primary thrust bearing206, a centering bearing208, and a power take off shaft210fixedly mounted to an underside of the thrust plate200along a thrust axis212which intersects the central longitudinal axis42. The thrust plate200revolves in a thrust plane around the thrust axis212and is supported against the thrust housing plate204by the primary thrust bearing206which is positioned against a flange214extending from an underside of the thrust plate200. The centering bearing208is positioned around the power take off shaft210adjacent a flange216extending from an underside of the thrust plate200. The thrust plate200is tilted at a fixed oblique angle to a plane which is perpendicular to the central longitudinal axis42which is between 0° and 90° degrees. The synchronizing gear202or other synchronizing mechanism is positioned on the thrust axis212at the center of the thrust plate200for interfacing with the synchronizing gear53extending from the cylinder carriage52for transferring torque therethrough and for synchronizing the thrust plate200and cylinder bank assembly12in a one-to-one rotational relationship at the fixed oblique angle, which can be approximately 45° to maximize the long axis of the oval trajectory and hence the torque. Adjusting other parameters to maximize torque may result in an actual optimal range of the thrust plate angle between 35° and 75°. The thrust plate200supports the outer ends58of all the connecting rods56which are cardan joints with a preferable double universal joint or a spherical rotatable ball joint mounted thereto via retainers218. The thrust plate200directs the connecting rods56on a circular course in unison with the pistons54as the cylinder bank assembly12rotates. Since the thrust plate200is at an oblique angle to a plane perpendicular to the central longitudinal axis42and since the pistons54are linked to the thrust plate200by the connecting rods56, the pistons54are forced to travel between an up most position within the cylinder which is top dead center (TDC) and a down most position within the cylinder which is bottom dead center (BDC) as they rotate about the central longitudinal axis42. When the major axes of the cylinders are arranged parallel to the central longitudinal axis, then TDC is at 0° of thrust plate rotation and BDC is at 180° of thrust plate rotation. In this arrangement, at TDC the major cylinder axis, the connecting rod and the central longitudinal axis lie in the same plane. In this configuration, it is not practical to advance the thrust plate more than a few degrees because the rod will clash with the cylinder wall as the system rotates.

As evident fromFIGS. 1 and 2, increasing the oblique angle which the thrust plate200makes with the plane perpendicular to the central longitudinal axis42would cause the cubic displacement in the combustion chamber of the cylinder46to increase to a maximum defined by the stroke, which is the distance that the piston54travels within the cylinder46as the rotation of the cylinder bank assembly12advances from TDC to BDC, and which is defined by the radius of the circular trajectory of the centers of the outer ends58of the connecting rods56as they travel about thrust axis212. Since the pistons54are linked to the thrust plate200by connecting rods56, the bottom of the rods are thus made to follow a circular trajectory with respect to the thrust axis212. This circular trajectory forms an oval trajectory both with respect to a plane perpendicular to the central longitudinal axis42and with respect to a plane which is parallel to the central longitudinal axis42. As the cylinder bank assembly12rotates it becomes possible to cause the pistons54to effectively dwell near the top of its respective cylinder thereby increasing the heat and pressure forces acting on the pistons54and significantly improving the thermal efficiencies of combustion. As used herein, “dwell” refers to a substantially non-sinusoidal piston movement with respect to its corresponding cylinder and rotation of the output shaft. In particular, piston movement is substantially reduced at the top of the cylinder in spite of rotation of the output shaft. This allows combustion of the fuel/air mixture to occur when the volume of the cylinder above the piston is substantially constant, which improves thermal efficiency. Another potential advantage of the pistons54being linked to the thrust plate200in this way is that the dwell lessens the inertia of the pistons54as they reciprocate within the cylinder thereby, in effect, further increasing overall performance of the engine10.

Referring toFIGS. 22-25, it has been determined that there are many factors which can improve the thermodynamic and mechanical efficiency of the above described embodiment. These factors include but are not limited to (1) the diameter of the piston, (2) the number of cylinders, (3) the length of the stroke from TDC to BDC, (4) the radius of the cylinder bank, (5) the radius of the thrust plate, (6) the displacements or offsets301,302of the thrust plate axis from the central longitudinal axis (FIG. 23) in the directions along axes X and Z, (7) the angle of the thrust plate200with respect to the cylinder bank12(FIG. 24) and with respect to about the X, Y and Z axes, (8) the tilt of the major cylinder axis42(and hence the cylinders46) in both a pitch412and a yaw414(FIGS. 8 and 22) (two degrees of rotational freedom relative to the central longitudinal axis42), and (9) the advancement or retardation (i.e. angular rotational offset430) of the bottom ends of the connecting rods by rotating the thrust plate200about the thrust axis322in the thrust plane (FIGS. 8 and 25).

From a thermodynamic perspective useful work per cycle (W) is defined as follows:
W=pdV

where p is the instantaneous pressure in the combustion chamber and dV is the change in volume of the combustion chamber. Thus, it is desirable to for the piston to dwell (remain stationary or substantially stationary with respect to the cylinder wall) at the top of the cylinder while substantially all of the fuel burns to increase the pressure of the gases and then for the piston to move downward in the cylinder as quickly as possible to increase the dV. Thus, it is desirable to have a constant volume burn wherein 10% to 90% of the fuel is burned while the piston remains at the top of the cylinder and while the volume of the combustion chamber remains constant or substantially constant. Sophisticated thermodynamic modeling is necessary in order to calculate the pressures within the cylinder. However, it is estimated that a constant or substantially constant volume burn is accomplished when the piston dwells at the top of the cylinder for a crank angle interval of between 20-30 degrees. Thus, the above-mentioned9factors may be used to manipulate the piston position to create the desired dwell and increased pressure and then to move the piston away as quickly as possible to increase the dV of the combustion chamber. Because the pressures and temperatures resulting from a constant or substantially constant volume burn are so much higher than in a traditional reciprocating internal combustion engine, and because the burn rate is so much faster than a traditional internal combustion engine, it will be possible to run the air-fuel mixture much leaner than in a traditional internal combustion engine. Running lean extends the burn rate and effectively limits how lean an engine may run. Running lean on demand will therefore provide greater efficiency gains at the sacrifice of power density. Running lean may also alleviate any detonation problems resulting from the extremely high temperatures and pressures. Of course, it will also be possible to alleviate detonation issues by adjusting the piston motion to better control the temperature and pressure within the cylinders.

Referring to the free body diagram inFIG. 7, a detailed vector analysis may be employed to analyze the affect of these factors on the piston's position and the effective torque arm {right arrow over (M)}T, as the engine rotates over 360° in order to maximize the thermodynamic and mechanical advantage of the configuration. An effective torque arm, {right arrow over (M)}T, is calculated because the engine produces a torque arm along three axes, some positive and some negative, which must be resolved together. The higher the cumulative magnitude of the effective torque arm or moment, {right arrow over (M)}T, the higher the overall advantage of the configuration. It should be noted that the work (W) done at the piston from a thermodynamic perspective and from using the pdV equation is the same as the moment calculated at the output shaft using the following vector analysis. To obtain the moment about the thrust plate (i.e. the effective torque arm) the following equation is used:
{right arrow over (M)}T={right arrow over (D)}MA·{right arrow over (F)}R
Where,{right arrow over (M)}T=total moment about the torque plate{right arrow over (D)}MA=distance vector from the torque plate axis to the center of the outer end of the connecting rod{right arrow over (F)}R=force vector applied to the torque plate by the connecting rod

To obtain the distance vector, DMA, we calculate the distance in each of the x, y, and z directions between the center of the thrust plate and the point at which the rod axis intersects the thrust plate. For terminology purposes,RCP(x,y,z)=rod connection point, where the connecting rod axis intersects the torque plateTPC(x,y,z)=torque plate center
Written plainly,
{right arrow over (D)}MA=RCPx−TPCx, RCPy−TPCy, RCPz−TPCz

To obtain FRwe must identify the force in the cylinder FCthat is applied to the piston. Since both ends of the connecting rod are free to rotate, the connecting rod can only apply a force along the axis of its length. Because the connecting rod is at an angle, μ, to the piston's direction of travel, we divide FCby the cosine of μ to obtain FR. Or

To obtain μ we must define a vector that describes the direction of FR, but not the magnitude (since this is still unknown). The vector describing the length of the connecting rod, LR, does just this. LRis defined as
{right arrow over (L)}R=RCPx−PPx, RCPy−PPy, RCPz−PPz
Where,PP(x,y,z)=piston position (intersecting point of cylinder axis and connecting rod axis)

To obtain the angle between the two vectors LRand FC, divide the dot product of LRand FCby the multiplicative product of their two respective magnitudes as given in the equation below.

We can now obtain the moment MTwith our original equation; however this moment may not be in the same direction as the axis of rotation of our drive shaft. The moment about the drive shaft axis is called MS. This moment has a unit vector in its direction msthat is defined as,

We can also define this unit vector based on the known geometry of the engine (i.e. the orientation of the drive shaft with respect to the axis of the system). Therefore we can identify the angle, λ, between MTand msas

λ=M->T·m->SM->T
We will multiply MTby the cosine of λ to obtain MS.
{right arrow over (M)}S=COS(λ)·{right arrow over (M)}T

By analyzing the piston position and the effective torque arm {right arrow over (M)}Tor {right arrow over (M)}Ssuch as in a Microsoft Excel™ Spreadsheet it has been discovered that the most important factors for creating a dwell sufficient for a constant or substantially constant volume burn and then for increasing the mechanical advantage by having a fast moving piston are the cylinder tilts, the angle of the thrust plate with respect to the cylinder bank in three rotational degrees of freedom which includes its tilt with respect to two axes which are perpendicular to the central longitudinal axis and its rotational angular offset about an axis parallel to the central longitudinal axis and intersecting the thrust axis, the displacement or offset of the thrust plate axis from the central longitudinal axis (in one embodiment, such that they do not intersect), and the advancement/retardation (i.e angular rotational offset) of the thrust plate about the thrust axis. It must be understood that all of the factors are configured into the fabrication orientation of the cylinders, cylinder bank and thrust plate with respect to each other and they are not meant to be adjusted in any way whatsoever once they are designed into the engine. FIGS.8and22-25show these variables which are used to custom contour the piston motion to create a dwell for combustion and then to quickly move the piston down within the cylinder.

Referring toFIGS. 8 and 22, tilting the major cylinder axis370so that it is not parallel to the central longitudinal axis342, provides significant piston dwell and better aligns the connecting rod axis374with the thrust plate when maximum torque is delivered. The top end of each cylinder is tilted about a tilt point410on the major cylinder axis370nearest the bottom end of the cylinder in a direction away from the central longitudinal axis342, so that the major cylinder axis370has both a pitch angle412and a yaw angle414. The pitch angle412is the tilt of the top ends of the cylinders into or away from the direction of rotation of the cylinder bank and is measured as the angle between a first plane416which includes the central longitudinal axis342and the tilt point410, and a projection418of the major cylinder axis370onto a second plane420which is perpendicular to the first plane416and parallel to the center longitudinal axis342and which includes the tilt point410. The yaw angle414is the tilt of the top ends of the cylinders into or away from the central longitudinal axis and is measured as the angle between a line422formed by the intersection of the first plane416and the second plane420, and a projection424of the major cylinder axis370onto the first plane416. Generally, the yaw angle414brings the lower ends of the cylinders together, while causing the upper ends to spread apart from each other. The probabilistic ranges for both the pitch angle412and the yaw angle414are between 0° and 70° depending on the configuration and the other factors.

The thrust plate angle was discussed above with regard to increasing the displacement of the engine. Referring toFIG. 24, it should be noted that the thrust plate angle includes an X tilt angle305which is an angle measured in a plane perpendicular to the central longitudinal axis342and including the X and Z axes, a Z tilt angle307which is an angle measured in a plane perpendicular to the central longitudinal axis342and including the X and Z axes, and a Y rotation angle309which is an angle measured by rotating the thrust plate200about the Y axis which is parallel to the central longitudinal axis342. All three tilts (i.e. three rotational degrees of freedom) of the thrust plate can be used to affect the motion of the piston to create the dwell and to quickly move the piston after the dwell.

Referring toFIGS. 8 and 23, the displacements or offsets301and302of the thrust plate axis from the central longitudinal axis342results from moving the cylinder bank12and/or thrust plate200laterally with respect to each other (seeFIGS. 12 and 16) so that the thrust plate axis and central longitudinal axis do not intersect. In order to synchronize rotation speed of the cylinder carriage12with the thrust plate200when these two axes are offset, it becomes necessary to use a cardan-type gear set in the power take off assembly as described below with respect toFIG. 12. In combination with the tilting of the major cylinder axis, one or both of the offsets of the thrust plate axis from the central longitudinal axis has a dramatic effect on the piston motion to create the dwell and to quickly move the piston after the dwell.

Referring toFIGS. 8 and 25, the angle430of advancement/retardation of the thrust plate200is defined as the angular rotationally offset of the thrust plate200about the thrust axis and with respect to a reference point in the thrust plane (represented by the thrust plate200). The angle of advancement/retardation430is the angular differential between two lines in the thrust plane, wherein the first line432is between the thrust axis322and a reference point at the outer end of the connecting rod when the piston is in the up most position, and wherein the second line434is between the thrust axis322and the outer end of the connecting rod after the thrust plate200has been advanced or retarded about the thrust axis while the cylinder bank remains fixed. In the traditional sense, when the piston is at TDC, the major cylinder axis370is substantially aligned with the rod axis374. The idea behind the advancement/retardation angle is that the thrust plate is advanced in the direction of rotation or retarded in the opposite direction of rotation so that the rod axis374is advanced or retarded, respectively, from the major cylinder axis372by an angle, α. This is equivalent to advancing or retarding the cylinder bank so that the rod axis374is advanced or retarded from the cylinder axis370by the angle, α. The probabilistic range for the advancement/retardation angle α measured on the thrust plate is between 0° and 35° in either direction about the up most piston position. The surprising and unexpected effect of advancing/retarding the thrust plate with respect to the cylinder bank is that it increases the duration of the power stroke to be greater or less than 180° and changes the motion of the piston within the cylinder from TDC to BDC to enhance the dwell and quickly move the piston after the dwell. The duration of the power stoke is measured in degrees of rotation of the thrust plate in the thrust plane using the outer end of the connecting rod as the reference point as the piston moves from TDC where it is in the up most position within the cylinder to BDC where it is in the down most position within the cylinder. Depending on engine parameters and application it may be desirable to vary the duration of the intake and power strokes compared to the compression and exhaust strokes. More particularly, it may be more desirable to shorten the duration of the power stroke so that the piston moves faster after the substantially constant volume combustion which takes place during the dwell.

With regard to the other factors it is desirable to increase the diameter of the pistons as large as possible to provide optimal rod clearance as the system rotates and also to increases the cubic displacement of the engine and power density. Reducing the number of cylinders improves rod clearance issues and permits a shorter stroke engine, but this has to be balanced with having a smooth running engine. The stroke of the engine depends on its application and engine speed-in higher speed engines it is desirable to a have the stroke equal to the diameter of the piston (i.e. bore size) to reduce mean piston speed and associated ring losses. The diameter of the cylinder bank and thrust plate must be balanced with the other engine parameters to achieve the desired stroke.

It must be understood that while the mathematical analysis may yield an optimal configuration for the piston position, there are practical limitations in constructing the parts so that the rods neither clash with their own cylinder walls nor the adjacent rods or cylinders walls as the cylinder bank rotates over a full 360°. Thus, while the mathematical analysis provides guidance in determining which factors are most important for maximizing mechanical advantage, all of the factors must be adjusted to properly configure the cylinder bank with respect to the thrust plate for rod clearance. As a practical matter, rod clearances may be most easily determined using three-dimensional computer modeling software like SolidWorks™ by SolidWorks Corporation of Concord, Mass. Rod clearance issues can dramatically limit the ability to configure an engine. One counterintuitive method for achieving rod clearance is to increase piston diameter and cylinder diameter and to nest the lower ends of the cylinders as close as possible to each other. This has the desirable effect of increasing the displacement of the pistons while shortening the stroke, thereby improving the power density of the engine and reducing piston speed.

FIG. 9, is a top plan schematic of the cylinder bank12showing the cylinders46tilted with both a pitch angle and a yaw angle wherein the top ends of the cylinders46are spaced apart from each other.FIG. 10is a side view schematic of the cylinder bank12and thrust plate200showing the cylinder tilt and the nesting of the lower ends of the cylinders46.FIG. 11, is a bottom plan schematic of the cylinder bank12showing the tightest nesting position wherein a leading edge of the lower ends of each cylinder is touching the adjacent cylinders. Nesting the lower ends of the cylinders46in this manner allows the radius of the cylinder bank12to be at a minimum, thereby minimizing centripetal forces.

Referring toFIGS. 22-25, one embodiment of a five cylinder engine without the torque plate axis being offset from the central longitudinal axis (i.e. without the cardan-type joint) is described by the following specifications:

7.65inchesEffective rod length which is the length of therod from the center of the outer end joint to theintersection of the rod's axis and the cylinder'saxis2.04inchesRadius of the cylinder carriage circle from thecenter of rotation to the center of the cylinder3.06inchesRadius of the thrust plate from its center to thecenter of the outer end of the connecting rod4.675inchesDiameter of the piston0degreesAngle of the thrust plate with respect to the Zaxis in a plane perpendicular to the centrallongitudinal axis50Angle of the thrust plate with respect to the Xaxis in a plane perpendicular to the centrallongitudinal axis30Angle of the thrust plate with respect to the Yaxis in a plane perpendicular to the centrallongitudinal axis40degreesYaw angle5degreesPitch angle10degreesAdvancement angle of thrust plate with respect ocylinder bank0inchesOffset of the x coordinate of the center of thetop surface of the thrust plate0inchesOffset of z coordinate of the center of the topsurface of the thrust plate

Referring toFIG. 26, piston motion for this embodiment is illustrated at500, which shows a substantial dwell502and then a fast moving piston region504. In contrast, piston movement for a conventional crankshaft internal combustion engine is illustrated at506, which has substantially no dwell.

Another embodiment of a five cylinder engine with the torque plate axis being offset from the central longitudinal axis (i.e. with the cardan-type joint) is described by the following specifications:

7.65inchesEffective rod length which is the length of therod from the center of the outer end joint to theintersection of the rod's axis and the cylinder'saxis2.04inchesRadius of the cylinder carriage circle from thecenter of rotation to the center of the cylinder3.06inchesRadius of the thrust plate from its center to thecenter of the outer end of the connecting rod4.675inchesDiameter of the piston0degreesAngle of the thrust plate with respect to the Zaxis in a plane perpendicular to the centrallongitudinal axis50Angle of the thrust plate with respect to the Xaxis in a plane perpendicular to the centrallongitudinal axis−15Angle of the thrust plate with respect to the Yaxis in a plane perpendicular to the centrallongitudinal axis40degreesYaw angle5degreesPitch angle10degreesAdvancement angle of thrust plate with respect ocylinder bank1inchesOffset of the x coordinate of the center of thetop surface of the thrust plate0inchesOffset of z coordinate of the center of the topsurface of the thrust plate

Referring toFIG. 27, piston motion for this embodiment is illustrated at510, which shows a substantial dwell512and then a fast moving piston region514. In contrast, piston movement for a conventional crankshaft internal combustion engine is illustrated at506, which has substantially no dwell.

It must be understood that there are countless possible combinations of the design factors which can create any desired piston motion and detailed thermodynamic study is required to determine the most optimal configuration, with strong consideration given reducing the complexity of the engine while maintaining the desired piston motion and fast moving piston after the dwell.

Referring toFIG. 12, another embodiment of the power take off assembly314is illustrated in partial schematic form. In this embodiment the power take off assembly314includes a synchronizing member316operatively connected to the cylinder bank assembly312and the thrust plate320so that the cylinder bank assembly312and thrust plate320rotate at the same speed, and so that a center axis322of the thrust plate320is offset with respect to the central longitudinal axis342in a direction along both the x and y axes, which provides greater mechanical advantage and/or improved rod clearance. More specifically, the power take off assembly314includes a donut-shaped thrust plate330which revolves about the center axis322which is offset from and does not intersect the center longitudinal axis342, a power take off332, a cardan-type gear set334for synchronizing the thrust plate330to the cylinder bank assembly312, and a stationary thrust housing336for supporting the thrust plate330, the power take off332, and the cardan-type gear set334. The donut-shaped thrust plate330includes a central opening338, a synchronizing gear339set into an inner surface thereof, and an output gear340set into a peripheral surface thereof. The power take off332includes an output shaft344and a power transfer gear346synchronized to the output gear340of the thrust plate330for transferring power therefrom in a one to one ratio. It should be noted that the power transfer ratio can be adjusted to meet any particular application. The stationary thrust housing336includes a first bearing surface348for supporting the thrust plate330, a second bearing surface349for supporting the power take off332, and a stationary shaft350which extends up through the central opening338of the donut-shaped thrust plate330forming an offset axis360which intersects the center axis322of the thrust plate330and the central longitudinal axis342and which rotatably supports the cardan-type gear set334thereabout. The cardan-type gear set334includes a torque tube362rotatably mounted on bearings (not shown) about the stationary shaft350on the offset axis360, an upper synchronizing gear364which meshes with a synchronizing gear366on the underside of the cylinder carriage352, and a lower synchronizing gear368which meshes with the synchronizing thrust plate gear339. The center axis322of the thrust plate330is offset from the central longitudinal axis342to optimize the piston motion to create the dwell and to quickly move the piston after the dwell.

FIGS. 13-21illustrates features of other embodiments of rotating barrel type internal combustion engines having further aspects of the present invention. In the embodiment ofFIGS. 13-14, the engine10″ rotates about a stationary central support shaft175, which is fixedly attached to stationary support housing600. An outer cover is indicated at601. Thus, in this embodiment the bearings (not shown) are generally about the support shaft175and not on the periphery of the cylinder bank12as in the earlier exemplary embodiment. The central support shaft175permits a common exhaust manifold602with a flat exhaust seal at the bottom of the engine. The length and shape of the exhaust pipes604from the cylinders to the common exhaust manifold602can be adjusted to tune the exhaust gases for desired Helmholtz effect.

As illustrated inFIGS. 15 and 16, the common exhaust manifold602includes a stationary exhaust gas pickup610, a rotating plate612which is attached to the ends of the rotating pipes604, and a rotating seal (not shown) between the stationary exhaust gas pickup610and the rotating plate612. The exhaust gas pickup610includes a blowdown area620which receives the initial exhaust gases which are under the highest pressures, and a secondary exhaust chamber622which continues for the balance of the exhaust stroke. The exhaust gases from the blowdown area620feed directly into a common stationary tail pipe624through an opening while the exhaust gases from the secondary exhaust chamber622first move in the direction of the rotating exhaust plate612between a flow plate626and the rotating plate612and then loop back underneath the flow plate626to the blowdown area620where they flow into the tail pipe624. A venturi effect is thus created in the stationary exhaust pickup610between the blowdown area620and the secondary exhaust chamber622wherein the higher pressure blowdown gases from one cylinder pull the remnant gases from the preceding cylinder out the tail pipe624. The rotating seal is made from conventional material and is positioned between the rotating plate612and the stationary exhaust gas pickup610to prevent exhaust gases from leaking out and from leaking between the blowdown area620and the secondary exhaust pickup622. It may be desirable to back pressure the exhaust seal to make sure there is no exhaust gas leakage.

Although the subject matter has been described in language directed to specific environments, structural features and/or methodological acts, it is to be understood that the subject matter defined in the appended claims is not limited to the environments, specific features or acts described above as has been held by the courts. Rather, the environments, specific features and acts described above are disclosed as example forms of implementing the claims. In addition, workers skilled in the art will recognize that changes may be made in form and detail without departing from the spirit and scope of the inventive concepts described herein. For example, slight modifications to the structure of the present invention which has been described with respect to internal combustion engines, would permit the functioning principals of the design to be applied to two-cycle, diesel, steam and sterling cycle pumps and engines.