CA2366360A1 - Inverse peristaltic engine - Google Patents

Abstract

An inverse peristaltic engine system having a fixed principal chamber (1), substantially circular in cross section, housing a series of traveling combustion chambers. The inner and outer interior walls of the principal chamber are contoured so that the width of the principal chamber varies with respect to position. The combustion chambers consist of a cylinder (5) housi ng two opposing pistons (10). The connecting rod on each of the pistons has bee n modified to include two pairs of wheels (9) which contact the contoured surfaces of the principal chamber's inner walls. As the cylinders travel through the principal chamber, the pistons are forced to move toward one another and apart from one another as the space between the walls of the principal chamber varies during their travel. Further provided is a pluralit y of ports in the ceiling and floor of the principal chamber and port interfac es in the top and bottom sides of the cylinders, which together provide for the intake and exhausting of fuel. Additionally, several drive spokes (7) are attached to the ring of cylinders so that it may rotate the driveshaft at th e center of the engine as the cylinders travel in a circular path through the principal chamber.

Description

TITLE OF THE INVENTION
Inverse Peristaltic Engine INVENTORS: THOMAS, C. Russell, a US citizen of 7433 Birch Bend, Covington, LA

CROSS-REFERENCE TO RELATED APPLICATIONS
Priority is claimed from US Provisional Patent Application Serial Nos.
l0 60/125,798, filed 23 March 1999; 60/134,457, filed 17 May 1999; 60/141,166, filed 25 June 1999 and 60/147,584 filed 06 August 1999, all hereby incorporated herein by reference.
In the US, this is a continuation in part of US Patent Application Serial No.
09/150,315, filed 09 September 1998, hereby incorporated herein by reference.
STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR
DEVELOPMENT
Not applicable REFERENCE TO A "MICROFICHE APPENDIX"
Not applicable 2o BACKGROUND OF THE INVENTION
1. Field of the Invention The present invention relates to a novel concept of an internal combustion engine utilizing inverse peristaltic action to provide a significant amount of torque as a plurality of interconnected, mobile combustion chamber travel within a fixed principal chamber or along a contoured peristaltic track.

2. General Background of the Invention In the field of engine technology, there is a significant amount of art that has slowly evolved over the years from the first steam engines to the latest rotary engines.
The design of engines, whether they are conventional, rotary or other types, has almost always included a stationary chamber, either cylindrical, round, oblong or the like, inside of which the movement of a definable member such as a piston or rotor serves to drive the engine. The present invention has departed from traditional engine arrangements, by replacing stationary chambers with a number of traveling cylinders that drive the engine as they travel within a fixed principal chamber or along a contoured peristaltic track.

Additionally, the present invention has replaced the complicated and restrictive system of valves, found in nearly all engines, with a superior port system.
BRIEF SUMMARY OF THE INVENTION
The present invention introduces a new concept to the art of power driven engines.
What is provided is an Inverse Peristaltic Engine system having a fixed principal chamber, substantially circular in cross-section, housing a series of traveling combustion chambers. The inner and outer interior walls of the principal chamber are contoured so that the width of the principal chamber varies with respect to position. The combustion chambers consist of a cylinder housing two opposing pistons. The connecting rod on 1o each of the pistons has been modified to include two pairs of wheels which contact the contoured surfaces of the principal chamber's inner walls. As the cylinders travel through the principal chamber, the pistons are forced to move toward one another and apart from one another as the space between the walls of the principal chamber varies during their travel. Further provided is a plurality of ports in the ceiling and floor of the principal chamber and port interfaces in the top and bottom sides of the cylinders, which together provide for the intake and exhausting of fuel. Additionally, several drive spokes are attached to the ring of cylinders so that it may rotate the driveshaft at the center of the engine as the cylinders travel in a circular path through the principal chamber.
The cylinders described above may also be arranged parallel to the driveshaft and 2o may contain opposing or single pistons. In these types of configurations the pistons ride along one or two contoured peristaltic tracks that enable them to pass through their various cycles.
Additional features of the system allow the various embodiments of the invention, which can function as a Diesel or Otto-cycle internal combustion engine, to be modified to function as a steam engine, a pneumatic engine, a hydraulic engine, a positive displacement blower, a reciprocating pump, a one-piece integrated centrifugal pump/engine unit, or a one-piece integrated generator/engine unit. Other features provide a mode of lubrication, a mode of cooling, and a manner for sealing the port interfaces from the surrounding atmosphere and the various ports from any lubricating fluids that may be present within the principal chamber.
Therefore, it is a principal object of the present invention to provide an Inverse Peristaltic Engine that produces a significant amount of torque and incorporates significant flexibility for responding to the demands of various types of systems;

It is a further object of the present invention to provide an Inverse Peristaltic Engine that continues to pass through its four cycles (induction, compression, combustion, and exhaust), as a series of cylinders travel within a fixed principal chamber or along a contoured peristaltic track(s);
It is a further object of the present invention to provide an engine whose total displacement can be increased or decreased by adding more or less cylinders, or increasing or decreasing the displacement of the individual cylinders through a longer or shorter stroke or by using larger or smaller cylinders within the system; and It is a further object of the present invention to provide an engine wherein the to contoured inner and outer walls of the principal chamber or the surface profile of the peristaltic tracks) may be modified to vary the engine's torque, RPMs and the individual expansion or compression rate, stroke length, time duration and compression ratio of each of its four cycles by changing the angle of incline and decline before and after or before or after the restricted neck portions of the principal chamber and/or by changing the width of the principal chamber in the intake and compression areas and/or the combustion and exhaust areas (or by altering the surface of the peristaltic track in a respective manner), and to provide an engine in which the overall compression ratio may be continuously varied during its operation by raising and lowering the peristaltic track.
BRIEF DESCRIPTION OF THE DRAWINGS
2o For a further understanding of the nature of the present invention, the detailed description of the invention should be read in conjunction with the following drawings, wherein like reference numerals denote like elements, and wherein:
Figure 1 illustrates an overall view of the preferred embodiment of the engine with the top section removed;
Figure 2 illustrates a cross-section view of the block portion of the engine forming the principal chamber, the secondary chamber and the crankcase cavity;
Figure 3 illustrates a partial view of the end portion of the block with the top and bottom sections bolted and sealed together to form the principal chamber, which is shown housing a traveling cylinder and its two pistons;
Figure 4 illustrates a top partial view of the principal chamber with all other components of the engine removed, exposing the intake ports, the fuel injectors, the and exhaust ports;
Figure 5 illustrates a cross-section view of the traveling cylinder assembly housed within the principal chamber, depicting the relationship between the individual pistons and the contoured surfaces of the principal chamber's inner walls;
Figure 6 illustrates an isolated traveling cylinder as it passes through the cycles of induction, compression, combustion and exhaust;
Figure 7 illustrates an isolated view of the interconnected traveling cylinders with their interface seals, port seals, scrapers, and seal lubricating oil rollers;
Figure 8 illustrates an end cross-section view of the principal chamber wherein there is depicted a pair of fuel injectors within the top and bottom walls of the principal chamber;
1o Figure 9 illustrates two self lubricating metal plates 28 recessed into the block of the engine to eliminate the need for oil rollers;
Figure 10 illustrates the various oil galleries that carry high-pressure oil from the drive shaft to the ring of traveling cylinders;
Figure 11 illustrates the oil galleries and the sliding oil grooves that lubricate the pistons and their wheels;
Figure 12 illustrates how intake flow and swirling can be maximized by offsetting the ports and port interfaces to create a vortex within the cylinder during the intake and combustion cycles;
Figure 13 illustrates a method that may be used to further increase the 2o compression ratio of the engine;
Figure 14 illustrates a way to prevent the pistons from attempting to twist on their axes;
Figure 15 illustrates an additional but less favorable method of the above;
Figure 16 illustrates a second embodiment of the present invention that eliminates the need for drive spokes through the utilization of a pair of large conic gears engaged to the ring of traveling cylinders;
Figure 17 illustrates a third embodiment of the engine that decreases friction and wear within the cylinders by eliminating the horizontal force exerted on the cylinders' interior walls;
Figure 18 illustrates a hinged oil tube that bridges the gap between the protrusions on the ring of cylinders and the connecting rods to deliver high-pressure oil directly to the oil galleries within the connecting rods which lubricate the wheels;
Figure 19 illustrates a fifth embodiment of the engine that also eliminates friction and wear within the cylinders;
Figure 20 illustrates a sixth embodiment of the engine in which the ring of cylinders remains fixed while the principal chamber rotates;
Figure 21 illustrates a cross-section view of a portion of the fixed ring of cylinders wherein there is depicted an intake valve, a fuel injector and an exhaust valve;
Figure 22 illustrates an end view of a number of air passages cut through the ring of traveling cylinders;
Figure 23 illustrates a brushless, alternating current, integrated generator/engine unit in which the armature remains stationary and the only moving parts are the internal workings of the engine;
Figures 24A-24D illustrate four recommended armature and magnet arrangements to be used in the generator/engine unit;
Figure 25 illustrates a sixth embodiment of the engine in which the cylinders are oriented parallel to the driveshaft;
Figure 26 illustrates a parallel cylinder version of the engine that has been constructed at an ideal scale to hold one four-stroke peristaltic pattern repeat;
Figure 27 illustrates an enlarged plenum area in the throat of the intake and exhaust ports;
Figure 28 illustrates an embodiment in which scoops extend over the drilled 2o portions of the ring of cylinders to circulate air through the cooling system;
Figure 29 illustrates the generator/engine unit, first introduced in Figure 22, as applied to the parallel cylinder configuration;
Figure 30 illustrates an additional embodiment of the engine in which the traditional opposing piston configuration has been reduced to one piston per cylinder in an attempt to both simplify and further reduce the size of the engine;
Figures 31 and 32 illustrate a cooling system to be used in the single piston per cylinder version of the engine;
Figure 33 illustrates a third generator/engine unit as applied to the single piston per cylinder version of the engine;
3o Figure 34 illustrates the most preferable of three cooling systems adapted for the single piston per cylinder generator/engine unit;
Figure 35 illustrates the second most preferable cooling system for the single piston per cylinder version of the generator/engine unit;

Figure 36 illustrates the third cooling system for the single piston per cylinder version of the generator/engine;
Figure 37 illustrates an additional embodiment of the engine in which the engine employs a port distributor disk to allow the cylinders to remain stationary;
Figures 38A and 38B together illustrate a top and bottom view of the port distributor disk;
Figures 39A-39D illustrate the changing positions of the port distributor disk as the two cylinders in view pass through four cycles;
Figures 40A and 40B illustrate top and bottom views of a port distributor disk that 1o consist of three concentric sections;
Figure 41 illustrates an embodiment of the engine in which the connecting rods are fitted with conic wheels to reduce cornering wear;
Figure 42 illustrates a single piston per cylinder version of the engine utilizing an open engine cooling system;
Figures 43 and 44A-44C illustrate an open engine cooling system that eliminates the need for an exterior fan;
Figure 45 illustrates a single piston per cylinder generator/engine unit utilizing an open engine cooling system that cools both the engine and generator sections of unit;
Figure 46 illustrates a generator/engine unit that uses fins to facilitate the flow of 2o air through the cooling system;
Figures 47 and 48A-48B illustrate the original perpendicular cylinder version of the engine using an open engine cooling system;
Figure 49 illustrates an annular collector channel that captures and reburns any blow-by gasses that may manage to escape the interface or port area seals;
Figures 50 and 51 illustrate an embodiment of the engine that has been modified to function as an integrated centrifugal pump/engine unit;
Figure 52 illustrates a generator/engine unit that has retained its driveshaft so that it may simultaneously serve as a stationary engine and drive other equipment as it generates electricity;
Figure 53 illustrates a single piston per cylinder version of the engine with a modified bearing 64 arrangement that may be advantageous for certain applications;
Figure 54 illustrates a modified version of the oil-cooled embodiment of the engine, previously depicted in Figures 31 and 32, that includes a centrifugal filler tube and an air displacer tube to ensure that the reservoir areas remain filled with oil;
Figure 55 illustrates a single piston per cylinder version of the engine in which the height of the peristaltic track is controlled by hydraulic lifters, allowing the overall compression ratio of the engine to be continuously varied;
Figure 56 illustrates a second continuously variable compression ratio embodiment of the engine in which the height of the peristaltic track is controlled by a number of worm drives linked, synchronized, and driven by a chain;
Figure 57 illustrates a third type of variable compression ratio device which uses an annular screw drive mechanism to vary the height of the peristaltic track 61;
to Figure 58 illustrates a fourth continuously variable compression ratio embodiment of the engine in which the lifting device consists of two cam rings;
Figure 59 illustrates a hydraulically controlled continuously variable compression ratio device that may be ideal for extreme operating conditions or to simply obviate the need for a modified peristaltic track;
Figure 60 illustrates an opposing piston version of the engine with continuously variable compression ratio abilities;
Figure 61 illustrates a fixed cylinder embodiment of the engine that has been modified to incorporate a continuously variable compression ratio device;
Figure 62 illustrates an additional fixed cylinder embodiment of the engine which 2o employs a simplified version of the variable compression ratio device illustrated in the previous figure;
Figure 63 illustrates an embodiment of the engine in which spark plugs travel with the cylinders;
Figure 64 illustrates a porous ceramic plate recessed into the head of the engine which can be used in place of the oil rollers or wicks depicted earlier in this application;
Figure 65 illustrates an embodiment of the engine utilizing reinforced, semi-porous metal-matrix interface and port area seals;
Figures 66A-66B illustrate an embodiment that includes stationary ceramic blades to create swirl during the intake cycle, an embodiment that includes a honeycomb-type 3o regenerator to reuse heat from the exhaust, and a special type of spring used to keep the interface and port area seals pressed against the head of the engine;
Figure 67 illustrates a single piston per cylinder version of the engine with a modified bearing arrangement that allows the driveshaft to bear the entirety of the engine's rotating mass and suspend the ring of cylinders so that only the interface seals contact the block of the engine;
Figure 68 illustrates an embodiment of the engine that includes a spherical sealing surface inside the head of the engine;
Figures 69A and 69B illustrate the outer port area seal and the crankcase seal;
Figure 70 illustrates an embodiment of the engine that utilizes a port distributor globe to aspirate the engine;
Figure 71 illustrates an embodiment of the engine in which the seals are readily accessible;
1o Figures 72A-72C illustrate two different types of brush and contact ring ignition systems;
Figure 73 illustrates a valve-aspirated version of the engine;
Figure 74 illustrates a valve-aspirated version of the engine utilizing hydraulic lifters to obtain continuously variable compression ratios;
Figure 75 illustrates an embodiment of the engine that reduces the length of the peristaltic plate/stationary cylinder embodiments of the engine that employ continuously variable compression ratio abilities;
Figure 76 illustrates an additional embodiment that that reduces the length of the peristaltic plate/stationary cylinder embodiments of the engine that employ continuously 2o variable compression ratio abilities;
Figure 77 illustrates a two-cycle Diesel version of the engine that includes scavenging pistons;
Figure 78 illustrates another two-cycle version of the engine that uses conventional scavenging methods;
Figure 79 illustrates an opposed cylinder configuration that was derived from the two-cycle version of the engine depicted in Figure 77;
Figure 80 illustrates an embodiment of the engine that combines the time-tested sealing ability of poppet valves with the superior aspiration of a port distributor globe;
and 3o Figures 81A-81D illustrate an isolated traveling cylinder on a peristaltic track at different consecutive periods in time to illustrate the versatility of the peristaltic track and to further clarify the peristaltic process.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS OF THE
s INVENTION
Figures 1-81D, illustrate the preferred embodiment of the present invention.
In general, the figures will refer to an inverse peristaltic engine which includes a plurality of interconnected cylinders containing pistons which are engaged to a peristaltic track or plate. There would be included valves or ports for admitting and expelling fluids from the cylinders during operation, and a means for igniting the contents of the cylinders to power the engine. The valves are actuated by a rotating disk or wheel, while the ports are sealed against a spherical head area of the engine. The spherical head might include secondary seals and channels to control emissions. The overall compression ratio of the engine may be continuously varied, and the combustion of the engine is controlled by brush and contact ignition system. These and other features will be clearly illustrated in the figures and discussed below.
Figure 1, where the system is viewed with the top section of the engine removed, comprises the principal chamber 1 housed within the engine block 2. The engine block 2 is seen as a continuous circular housing having an outer wall 3 and an inner wall 4, both substantially circular in cross section. The principal chamber 1 extends uninterrupted throughout the entire circular housing and consists of a specific pattern of expanded and restricted portions that repeats itself throughout the entire length of the principal chamber 1. Housed within principal chamber 1 is a ring of interconnected traveling cylinders 5.
The ring of traveling cylinders transmits rotary motion to the driveshaft 6 through a plurality of drive spokes 7. Also illustrated in Figure 1 is a secondary chamber 8 which accommodates the outer wheels 9 of the pistons 10 contained within the traveling cylinders 5.
As illustrated in cross-section in Figure 2, the principal chamber l and the secondary chamber 8 are formed from a circular top section 11 bolted onto a circular lower section 12 with the central hub portion of the sections 13 securing the driveshaft 6 and the distal ends of the sections meeting at a common point 14. The two sections 11, 12 further define a crankcase cavity 15 which houses the drive spokes 7. A
continuous slit 16 is seen cut into the inner wall of the principal chamber. The continuous slit 16 is required to accommodate the width of the drive spokes 7, so they may penetrate the principal chamber 1, and the connecting rods 17 (see Figure 1) that attach the wheels 9 to the pistons 10, so they may penetrate the crankcase cavity. A secondary continuous slit 18 is cut into the outer wall of the principal chamber 1. This slit allows the outer connecting rods 17 (see Figure 1 ) to penetrate the secondary chamber 8 to attach to their respective wheels 9.
In Figure 3 there is illustrated in isolated view the end portion of the upper section 11 firmly secured to the end portion of the lower section 12 by a number of bolts 19. In this particular arrangement, it should be noted that rather than having a flat connection 14, there is provided a series of interlocking teeth 20, which allow the upper and lower sections 11, 12 to be more easily aligned and to engage more securely to one another.
Although not visible in this figure, there are also teeth, perpendicular to the teeth shown, that radiate from the center of the engine. Once again, there is seen clearly the principal to chamber 1 with the continuous slit 16 in its inner wall and the secondary continuous slit 18 in its outer wall. Also seen is the secondary chamber 8 that accommodates the outer wheels 9 of the connecting rods 17. Seen in greater detail than in the previous figures is a pair of opposing pistons 10 housed within a traveling cylinder 5. Each of the pistons has a connecting rod 17, each of which holds two pairs of wheels 9. The outer pairs of wheels are responsible for separating the pistons 10 during the induction cycle and the occasional misfire while the inner pairs of wheels move the pistons together during the compression and exhaust cycles. Also seen in Figure 3 are two ports 21 one cut through the ceiling and another cut through the floor of the principal chamber. During the induction cycle, the two ports 21 line up with the port interfaces 22 which are cut through 2o the top and bottom sides of the cylinder 5, to allow fresh air to be admitted into the cylinder.
In Figure 4 a portion of the principal chamber 1 is viewed with all of the moving components of the engine removed. As stated earlier, the contoured walls of the principal chamber 1 form a pattern of expanded and restricted areas that is repeated throughout the entire length of the principal chamber 1. Returning to Figure 4, a plurality of ports 21, 24 and fuel injectors 23 are seen in the floor of the principal chamber 1. First, there is seen an intake port 21, soon followed by a fuel injector 23, and finally an exhaust port 24.
This pattern of port 21 fuel injector 23 port 24 is repeated throughout the entire length of the principal chamber 1. Although not visible in this illustration, intake ports 21, fuel injectors 23 and exhaust ports 24 are also found in the ceiling of the principal chamber 1 in the same positions as shown on the floor, (see Figures 3 and 8).
Figure 5 illustrates a cross-section view of a series of traveling cylinders 5 moving through the expanded and restricted areas of the principal chamber 1. Also illustrated are the partial features of two inlet ports 21, a fuel injector 23, and two exhaust ports 24 as they relate to the various cylinders 5 within the system.
For a full understanding of the main operational concept of the Inverse Peristaltic Engine, reference is given to Figures 6A through 6J where a single isolated traveling cylinder is illustrated as it passes through the cycles of induction, compression, combustion and exhaust. Although they would not normally be visible in this particular type of cross-section view, the port interfaces have been included in order to illustrate their positions in relation to the various ports and fuel injectors. While these figures depict only a single traveling cylinder, it should be understood that each cylinder within the system is constantly undergoing these same processes pursuant to its own timetable.
In Figure 6A, an isolated traveling cylinder 5 is depicted within one of the narrow necks of the principal chamber 1. At the cylinder's present position, the pistons 10 have come together fully, and the port interface 22 is just over front end of an intake port 21.
In 6B, the cylinder 5 has left the narrow neck of the principal chamber 1 and is completely over the intake port 21. As the pistons 10 separate, they form a vacuum within the cylinder 5, which draws in fresh air through the intake port 21. In 6C, the cylinder 5 has entered a wide neck in the principal chamber 1 and the pistons 10 have fully separated and completely filled the cylinder 5 with fresh air. (Notice that the inner and outer walls of the wide neck of the principal chamber 1 are of sufficient length so that the port interface 22 may fully clear the intake port 21 before the cylinder 5 begins its compression cycle.) In 6D, the cylinder 5 is entering a narrow neck of the principal chamber 1, forcing the pistons 10 together and hence compressing the air. In 6E, the cylinder 5 has completely entered the narrow neck of the principal chamber l and fully compressed the air. At this point the port interface 22 is lined up with the second fuel injector 23. The fuel injector 23 will now inject a fine mist of diesel fuel which will ignite upon contact with the hot, compressed air, thus commencing the combustion cycle.
(The first fuel injector 23 is only used when the engine is run at high RPMs and there is less time to inject fuel into the cylinders 5. Under these circumstances, the first fuel injector 23 switches on to work in unison with the second injector 23 in order to provide the additional injection time needed to deliver the proper amount of fuel.) In 6F, the expanding gases push the pistons 10 apart, forcing the cylinder 5 to roll into a wide neck of the principal chamber 1. In 6G, the cylinder 5 has traveled into the wide neck of the principal chamber 1, thus ending the combustion cycle. (Notice that the exhaust port 24 has been placed where it will not interfere with sealing during the combustion process.) In 6H, the port interface 22 is over an exhaust port 24 and the pistons are being forced together as the cylinder travels into a narrow neck of the principal chamber 1. As the pistons 10 approach one another, they create a high pressure within the cylinder 5, which expels the exhaust fumes left over from the combustion cycle. In 6I, the cylinder 5 has fully exhausted and is now in the narrow neck of the principal chamber 1.
(Notice that the exhaust port 24 and intake port 21 are appropriately distanced to allow the optimal level of port overlap suited for the demands placed on the system.) In 6J, the cylinder 5 is over a second intake port 21. Once again, the cylinder 5 will refuel as it enters another 1 o wide neck of the principal chamber l and begins to repeat the cycles.
Reference is now given to Figure 7 where a series of traveling cylinders 5 is illustrated moving through the expanded and restricted areas of the principal chamber 1 with the drive spokes 7 passing through the continuous slit 16 in the principal chamber's inner wall. Surrounding the port interfaces 22 is a pair of concentric, square interface seals 25. The interface seals 25 seal the port interfaces 22 from the surrounding atmosphere during the engine's compression and combustion cycles. The square shape of the interface seals allows each of the seals' four sides to wear evenly and causes the seals to wear an even trench into the ceiling and floor of the principal chamber 1.
These elements combined ensure that the interface seals 25 maintain good, level contact with the 2o ceiling and floor of the principal chamber 1 over a long period of time.
The square shape of the interface seals 25 also allows the engine to have the ability to function equally while rotating clockwise or counterclockwise. Surrounding the interface seals 25 are the oil seals 26. The oil seals 26 prevent any lubricating fluids that may be present within the principal chamber 1 from seeping into the intake and exhaust ports 21,24 by keeping the area on the ceiling and floor of the principal chamber that surrounds the ports 21,24 free from oil. Because the interface seals 25, which need lubrication to promote good sealing and to prevent premature wear, fall within this area, either an oil roller 27 or a high density cloth wick is placed before each pair of interface seals to provide the needed lubrication. A thick coating of oil in this area would defeat the purpose of the oil seals 26, so the oil rollers 27 function on the same concept as that of a ball point pen, rolling a thin coating of oil over the contacting surfaces on the ceiling and floor of principal chamber 1 without dripping oil into the ports 21,24. If wicks are used in place of rollers, they will be kept slightly damp so that they too may only apply a thin coating of oil. Also found in the area between the oil seals 26 are four scrapers 28. The scrapers 28 even out the wear area just off the inner and outer tips of the smaller interface seals 25. At this point the wear areas of the smaller and larger interface seals 25 do not overlap, so additional wear is necessary to maintain a single level wear trench.
Additionally, the angle and placement of the oil seals 26 are such that their wear areas on the surface of the principal chamber 1 will be of equal depth and will be contiguous to the wear areas of the interface seals 25, again maintaining a single level wear trench. It also should be noted that the length of the oil rollers 27 matches the width of the wear trench, so that over time, the rollers 27 will remain in good contact with the wear prone areas.
to To further illustrate the position of the fuel injectors 23, reference is given to Figure 8 where there is seen an end view of the upper and lower sections 11, 12 forming the principal chamber 1 with the continuous slit 16 in its inner wall and the secondary continuous slit 18 in its outer wall. Also seen is the secondary chamber 8 that accommodates the outer wheels 9 of the various connecting rods 17. In this particular view, there are two fuel injectors 23, one in the upper section 11 and the other in the lower section 12. It should be noted that the fuel injectors 23 are flat-faced so that they do not extend into the principal chamber 1. The various seals 25,26 can also be seen here pressed against the ceiling and floor of the principal chamber 1 by small springs concealed beneath them. The seals 25,26 and the scrapers 28 are similar in construction 2o to piston rings in that they fit into recessed grooves cut into the top and bottom of the ring of traveling cylinders 5.
Figure 9 illustrates two metal plates 28 recessed into the block of the engine. The two plates 28 consist of a metal infused with dry lubricants to allow the oil rollers 27 or wicks, previously depicted in Figure 7, to be eliminated. By threading the holes in the plates that the ends of the fuel injectors 23 pass through, it is possible to use the fuel injectors 23 to secure the plates 28 to the block.
Figure 10 is an X-ray view of the oil galleries 30 that carry high-pressure oil to the various moving components of the engine. The oil galleries 30 originate in the driveshaft 6 and travel through the drive spokes 7 to connect with the four main galleries 31 in the 3o ring of cylinders 5. From here the galleries 30 branch off to bring lubrication to where it is needed within the engine.
Figures 11 A and 11 B illustrate the pistons 10 with a sliding groove 32 that connects the oil galleries 30 within the pistons 10 and connecting rods 17 to the main galleries 31 within the ring of cylinders. As can be seen through a comparison of Figure 11A and Figure 11B, the length of the groove 32 allows this connection to be maintained regardless of the pistons' 10 position the within the cylinders 5 to provide continuous lubrication for the pistons 10 and wheels 9.
Figure 12 illustrates how intake flow and swirling can be maximized by offsetting the ports 21,24 and port interfaces 22 to create a vortex within the cylinders 5 during the intake and combustion cycles. During the intake cycle, the vortex formed within the cylinders 5 minimizes turbulence, therefore reducing intake resistance and allowing a larger volume of air to be admitted into the cylinders 5 during the allotted intake time. In to the combustion cycle, the swirling effect of the vortex produced during the early stages of combustion promotes good mixing of the fuel and air and circulates the flame to ensure a clean and complete burn.
In Figure 13 a ridge 33 protrudes from the ceiling and floor of the principal chamber 1 to slide within a trench 34 cut into the top and bottom of the ring of cylinders 5. This configuration minimizes the depth of the port interfaces 22 without reducing the thickness of the walls of the cylinders 5, thus allowing for a higher compression ratio without sacrificing the integrity of the ring of cylinders 5.
Figures 14 and 15 illustrate methods of preventing the various pistons 10 from attempting to rotate on their axes. In figure 14, the primary and secondary continuous 2o slits 16,18 have been narrowed and the connecting rods 17 have been fitted with a flat area 35. The flat areas 35 slide securely within the continuous slits 16,18, preventing the pistons 10 from attempting to rotate. In figure 15 a pair of fins 36 has been included on each of the pistons 10. These fms 36 slide within a pair of grooves 37 cut into the inner walls of the cylinders 5, again preventing the pistons 10 from attempting to rotate.
Rotation of the pistons is highly unlikely to pose a problem in the Inverse Peristaltic Engine for two reasons: the first is that there is no measurable force that motivates them to do so and the second is that the only times that rotation of the pistons is physically possible is when the wheels on the connecting rods pass through areas of the principal chamber where the inner and outer walls have no slope. These areas of no slope occur only at the peaks and valleys of the peristaltic pattern and span such a minute distance that, in most cases, the wheels on the connecting rods will be too large to contact them exclusively. Bearing this in mind, it should be reasonable to assume that in a normal peristaltic configuration no course of action will be necessary to prevent the pistons from rotating.
Figure 16 illustrates a second embodiment of the present invention that eliminates the need for drive spokes 7 through the utilization of a pair of large conic gears 38 engaged to a toothed surface 39 on the ring of traveling cylinders 5. In exchange for torque, this configuration significantly increases both the RPM output of the engine and the structural integrity of the principal chamber 1.
Figure 17 illustrates a third embodiment of the engine that nearly eliminates all friction and wear within the cylinders 5 by eliminating the horizontal force exerted on the to cylinders' S interior walls. In this embodiment, the ring of cylinders 5 includes protrusions 40 that penetrate the inner and outer continuous slits 16,18 to receive the horizontal force directly from the connecting rods 17. Each of the protrusions 40 holds two wheels 41 which grip the sides of the connecting rods 17 and roll back and fourth as the connecting rods 17 reciprocate with the pistons. In addition to eliminating wear within the cylinders 5, this embodiment also allows shorter pistons 10 to be used within the engine, thus significantly reducing the engine's reciprocating mass.
Although not illustrated in Figure 17, the attachment between the pistons 10 and connecting rods 17 may be hinged to ensure that any imprecision where the wheels 41 of the protrusions 40 grip the connecting rods 17 does not translate into a minor horizontal force still being 2o transferred to the walls of the cylinders 5.
When short pistons 10 are used in the third embodiment of the engine, it becomes difficult to supply oil to the wheels 9 on the connecting rods 18 using the original method of lubrication (see Figures 11 A and 11 B). Figure 18 illustrates a hinged oil tube 42 that bridges the gap between the protrusions 40 on the ring of cylinders 5 and the connecting rods to deliver high-pressure oil directly to the oil galleries 30 within the connecting rods 17 which lubricate the wheels 9. The oil tubes 42 receive their oil supply from oil galleries that branch off from the main galleries 31 that run within the ring of cylinders S
(see Figures 10, 1 lA and 11B). While this method of lubrication is quite unconventional, the length of the oil tubes 42 is such that they need pivot only slightly as they reciprocate 3o with the connecting rods 17, thus reducing any wear in their joints to an absolute minimum and ensuring exceptional reliability. If deemed more desirable, it is also possible to use a flexible hose in place of the hinged tube shown 42.
Figure 19 illustrates a fourth embodiment of the engine which also decreases friction and wear within the cylinders 5. In this embodiment, the horizontal force is received by a pair of bars 43 attached to the connecting rods 17. The bars 43 connect the connecting rods 17 to a plate 44 at the end of the drive spoke 7. The connections 45 between the connecting rods 17 and the bars 43 and the bars 43 and the plate 44 are hinged in order to compensate for the vertical travel of the pistons 10. Due to the fact that the cylinders 5 are no longer linked together, much of the system's original stability has been lost. To compensate for this lack of stability, a pair of tracks 46 has been included in the ceiling and floor of the principal 1 chamber and a pair of extensions 47 has been included on the top and bottom of the cylinder 5 and the top and bottom of the plate 44 at 1 o the end of the drive spoke 7. The extensions 47 slide within the tracks 46 as the linkage travels through the principal chamber 1 to keep the cylinder 5 properly aligned and give additional support to the drive spoke 7.
Figure 20 illustrates a fifth embodiment of the engine. The fifth embodiment operates under the same general theory as the previous embodiments; however, in this configuration the ring of cylinders 5 remains fixed while the principal chamber 1 rotates.
To allow the principal chamber 1 to rotate, a primary block 48 has been designed to encompass and support the preexisting block 2, which, being the frame of the principal chamber 1, must now have the ability to rotate. In addition, the drive spokes have been removed from the ring of cylinders 5 and attached to the block 2, and, because the 2o cylinders 5 are no longer traveling, the ports 21,24 have been replaced with a more conventional valve setup. The valves 49,50 are actuated by a contoured wall 51, similar to the contoured walls of the principal chamber 1, that extends from the traveling block 2 of the engine. This valve setup also includes a rocker 52 and a modified pushrod 53 fitted with a wheel 54. As the wheel 54 rolls over the peaks and valleys of the contoured wall 51, the pushrod 53 tips the rocker 52, which in turn actuates the valve 49.
Although efficient and durable, this method of controlling the various valves 49,50 could easily be substituted with a valve setup similar to that used in radial aircraft engines.
To further enunciate the properties of the preceding embodiment, Figure 21 illustrates a cross-section view of a portion of the fixed ring of cylinders 5 wherein there 3o is depicted an intake valve 49, a fuel injector 23 and an exhaust valve 50.
Because this engine utilizes opposing pistons 10, the valves 49,50 and fuel injectors 23 are placed on the side of the cylinders 5.
Figure 22 depicts the primary section of the Inverse Peristaltic Engine's cooling system. Illustrated is an end view of a number of air passages 55 cut through the ring of traveling cylinders 5. In this embodiment several holes are drilled into the outer wall of the secondary chamber 8 and the crankcase cavity 15 is vented by a high power fan.
When the fan is activated, it forms a vacuum within the crankcase 15 which draws cool air through the holes in the secondary chamber 8 and into the outer portion of the principal chamber 1. From here the air travels through the air passages 55 and in the process removes excess heat from the ring of cylinders 5. The hot air then enters the crankcase cavity 15 where it is blown out by the fan and routed through an intercooler before reentering the system. For fewer moving parts, it may be possible to eliminate the 1 o fan apparatus entirely by modifying the rotating internal workings of the engine to function as a centrifugal blower. In this embodiment, air would enter at the crankcase and be drawn outward through the cooling system by centrifugal force.
In Figure 23, the embodiment of the Inverse Peristaltic Engine seen in Figure has been further modified to function as a brushless, alternating current, integrated generator/engine unit in which the armature 56 remains stationary and the only moving parts are the internal workings of the engine. As with the embodiment in Figure 16, the crankcase cavity 15 and its contents have been removed; however, in Figure 23, the secondary chamber 8 has been enlarged to house the generator portion of the unit. In the generator portion of the unit, a number of magnets 57 can be seen suspended between the 2o armatures 56 by extensions 58 attached to the ring of cylinders 5. (The magnets have been extended outward from the ring in order to attain higher linear velocities.) In addition to the enlarged secondary chamber 8, a non-conducting gasket 59 has been placed between the outer ends of the top and bottom sections 11,12 of the engine to prevent the possible formation of an electric current within the block.
According to the layout presented in Figure 22, a unit containing sixty magnets 57 and sixty armature pairs 56 (the armature shown in Figure 23 is one pair) would produce an alternating current at a frequency of sixty hertz with the engine running at one-hundred and twenty RPMs. As will soon be illustrated in Figures 24A-24D, the generator/engine unit can be modified to produce a direct current by reconfiguring the polarity of the magnets 57 and rewiring the 3o armature 56. In addition, it may be possible to configure the magnets 57 and the block of the engine so that the armature 56 may be placed externally. The generator/engine unit may also make use of a more conventional arrangement in which the magnets 57 remain fixed to the block of the engine while the armature 56 rotates; however, this arrangement will require brushes to transmit power from the armature 56.
Figures 24A-24D illustrate four recommended armature 56 and magnet 57 arrangements to be used in the generator/engine unit.
Figure 24A illustrates the alternating current arrangement used in Figure 22.
In this arrangement the polarity of the magnets 57 alternates from one magnet to another.
Figure 24B illustrates a direct current arrangement in which all of the magnets 57 have the same polarity. Figure 24C illustrates a simplified alternating current arrangement while Figure 24D illustrates a simplified direct current arrangement.
Figure 25 illustrates a sixth embodiment of the engine. In this embodiment the cylinders 5 are oriented parallel to the driveshaft 6 instead of perpendicular to the driveshaft as in the original embodiments of the engine. Seen in new positions, shapes and sizes are the ring of cylinders 5, the crankcase cavity 15, driveshaft 6 and drive spokes 7, the two secondary chambers 8, the fuel injectors 23, ports 21,24 and port interfaces 22, with an exhaust manifold 60 referenced for the first time. In this cylinder configuration the diameter of the engine no longer affects the slope of the peristaltic walls, thus allowing the size of the engine to be drastically reduced. To allow the engine to be easily assembled, the connecting rod assembly 17 on each of the pistons 10 has been adapted to carry only one pair of wheels 9 and the peristaltic walls, hereinafter referred to as the peristaltic track 61, have been modified accordingly.
2o To further emphasize the size reduction capabilities of the parallel cylinder configuration, the engine in Figure 26 has been constructed at an ideal scale to hold one four-stroke peristaltic pattern repeat. This engine, as well as those depicted in the following figures, will function best with four to seven cylinders 5, depending on the diameter of the engine.
The Engine in Figure 26 can be further reduced in size by configuring it to function as a two cycle engine; however, this is somewhat undesirable because a two-cycle pattern robs the peristaltic configuration of much of its defining flexibility that allows the expansion or compression rate, stroke length, time duration and compression ratio of each the engine's four cycles to be fully adjustable and to vary from one cycle to 3o another (See Efficiency). A two-cycle pattern could, however, be configured with dwell time to provide the engine with superior scavenging capabilities.
Seen in partial view in the preceding figure and clearly in the cross-section view in Figure 27, is an enlarged plenum area 62 in the throat of the intake and exhaust ports 21,24. If the ports 21,24 were cut completely through the block for the entire span of the intake and exhaust strokes in an engine having only one four-stroke peristaltic pattern repeat, the integrity of the block would be severely compromised. The plenum areas 62 allow the ports 21,24 to be cut only partially through the block for the majority of the span of the intake and exhaust strokes to preserve the strength of the block.
Also seen in Figure 26 is a multitude of small holes 55 drilled through the ring of cylinders 5. The holes 55, which could also be substituted for slits, are derived from the embodiment depicted in Figure 22 and function as the primary section of the engine's cooling system.
In the embodiment in Figure 26, cool air is first drawn into one of the engines two 1o secondary chambers 8. From here the air proceeds to travel through the holes 55 in the ring of cylinders 5 where it removes excess heat from the cylinders 5 before entering the second secondary chamber 8 as hot air. The hot air is then drawn out of the second secondary chamber 8 by a high power fan and, as was done in Figure 22, routed through an intercooler before it reenters the system.
To eliminate the need for a fan assembly, the embodiment illustrated in Figure includes air scoops 63 that extend over the drilled portions 55 of the ring of cylinders 5.
As the ring of cylinders 5 rotates, the scoops 63 at the top portion of the ring form a positive pressure above the holes 55 while the scoops at the bottom of the ring form a negative pressure below the holes. This pressure difference forces air in the top 2o secondary chamber 8 to pass through the holes 55 and to continue to circulate through the system.
Figure 29 illustrates the generator/engine unit, first introduced in Figure 23, as applied to the parallel cylinder configuration.
Figure 30 illustrates an additional embodiment of the engine in which the traditional opposing piston configuration has been reduced to one piston per cylinder in an attempt to both simplify and further reduce the size of the engine. In this configuration the ports 21,24 have been moved to the top of the engine where they interface directly with the open top ends of the cylinders 5. Because this engine does not have opposing pistons 10, an upward force is now exerted on the ring of cylinders 5. To counteract any 3o negative effects of this force, a number of bearings 64 has been placed between the ring of cylinders 5 and the top inside wall of the engine to reduce friction between the contacting surfaces.
Because of the new port placement in the single piston per cylinder version of the engine, it was necessary to devise a new type of cooling system. Figures 31 and 32 illustrate the various features of this system, which utilizes oil as a cooling fluid. In Figure 32 the ring of cylinders 5 is seen isolated from the engine. Except for two disc sections, all of the portions of the ring that lie between the cylinders 5 have been removed, leaving an open area surrounding the cylinders which, when placed within the block of the engine, serves as a reservoir 65 for the cooling oil. On the top disk section of the ring, four port interfaces 22 are seen positioned above their respective cylinders 5. At the center of the disk section there is seen a fifth hole 66, the purpose of which, as will soon be discussed in detail, is to introduce cooling oil to the reservoir area 65 between the l0 cylinders 5. Now turning to Figure 31, there is seen a cross section of the engine illustrating three oil passages 67 cut through the block: one cut through the top of the block to allow oil to enter the reservoir and two cut through the side of the block to allow oil to exit. When the engine is activated, the cylinders 5 act as a vane and cause the oil that surrounds them to rotate with the ring of cylinders 5. As the oil's rotational velocity increases, centrifugal force creates a pressure difference between the inner and outer portions of the oil's rotating mass. This pressure difference draws cool oil through the hole 66 in the upper disk section of the ring of cylinders 5 and into the reservoir 65.
Upon entering the reservoir 65, the oil flows outward, removing heat from the cylinders 5 before eventually exiting as hot oil through the two passages 67 in the side of the block.
2o The hot oil then passes through a radiator 68 where it is cooled before reentering the system. Although oil provides a more thorough heat exchange, this system may also be adapted to use air as a cooling fluid.
Figure 33 illustrates a third generator/engine unit as applied to the single piston per cylinder configuration of the engine.
Figure 34 illustrates the most preferable of three cooling systems adapted for the single piston per cylinder generator/engine unit. In Figure 34, the ring of cylinders 5 includes reservoir areas 65 between each of its cylinders, which, as in Figures 31 and 32, are supplied with a continuous flow of cool oil. In addition, the block of the engine has been modified to include two annular groves 69 that form a continuous interface with the reservoir openings 70 to allow cooling oil to enter and exit each reservoir 65. As the engine/generator unit begins to rotate, centrifugal force creates a pressure difference between the inner and outer reservoir openings 70. This pressure difference draws cool oil from the inner annular grove 69 into the top portion of the reservoirs 65.
Upon entrance, the oil commences to flow down through the reservoirs 65, removing heat from the cylinders 5 before exiting into the outer annular 69 grove as hot oil.
From the outer annular grove, the hot oil then passes through one of two galleries 71 and one of two radiators 68 before eventually returning to the reservoirs 65 as cool oil.
Figure 35 illustrates the second most preferable cooling system for the single piston per cylinder version of the generator/engine unit. This cooling system operates under the same concept as the previous air-cooled systems; however, in this system air must enter from the side of the ring of cylinders 5 so as not to interfere with the port areas at the top of the engine. As in the previous air-cooled systems, cool air is drawn through to holes or slits 55 in the areas between the cylinders to remove excess heat from the engine (refer to Figures 21 and 26 and their respective text). Figure 35 also illustrates, for the first time in this application, the fan 72 and the intercooler 73.
The third cooling system for the single piston per cylinder version of the generator/engine unit is illustrated in Figure 36. Similar to the system introduced in Figure 34, this cooling system also uses oil as a cooling fluid, however, it does not make use of centrifugal force. Instead it uses a high volume oil pump (not shown) to circulate oil through the system. Although the requirement for an oil pump is not desirable, this system will be advantageous under extreme operating conditions because oil exiting the reservoirs 65 is directed to provide supplementary lubrication to the peristaltic track 61.
2o Figure 37 illustrates an additional embodiment of the engine in which the cylinders 5 remain stationary. In some applications where the engine is required to make sudden extreme changes in running speed, it will be desirable to have a low flyweight. In the embodiment in Figure 37 the peristaltic track 61 has been replaced with a peristaltic plate 74 which rotates with the driveshaft 6 while the cylinders 5 remain stationary. This configuration eliminates the ring of cylinders 5 which was responsible for the majority of the previous engines' rotational inertia. Although the engine now utilizes stationary combustion chambers, the simplicity, reliability and superior aspiration of a port system were able to be salvaged through the use of a port distributor disk 75. The port distributor disk 75 is attached directly to the top of the driveshaft 6 and has slits or port bridges 76 that connect the intake and exhaust ports 21, 24 to the cylinders 5 during the appropriate cycles. In order to properly time the intake and exhaust cycles, the port bridges 76 travel directly above the intake and exhaust stroke portions of the peristaltic plate 74. In Figure 37 the port bridge 76 is seen connecting the engine's right cylinder 5 to its respective exhaust port 24 while the left cylinder 5 remains blocked off. Without a traveling ring of cylinders 5, the cylinders 5 can no longer share common ports 21,24, so in this embodiment each cylinder 5 has its own intake and exhaust ports 21,24. Figures 38A and 38B together illustrate a top and bottom view of the port distributor disk 75.
38A is a top view of the disk 75, illustrating the portions of the intake 77 and exhaust bridges 76 that interface with the ports 21,24, while 38B is a bottom view of the disk 75, illustrating the portions of the intake 77 and exhaust 76 bridges that interface with the cylinders 5.
Figures 39A-39D illustrate the changing positions of the port distributor disk 75 as to the two cylinders 5 in view pass through four cycles. Each new figure represents one quarter turn of the driveshaft 6.
Reference is now given to Figure 39A. At this moment the port distributor disk 75 has connected the left cylinder 5 to its respective exhaust port 24, allowing it to exhaust while the right cylinder 5 remains blocked off during its compression cycle.
Advancing one-quarter turn to Figure 39B, the port distributor disk 75 has now connected the left cylinder 5 to its respective intake port 21 during its intake cycle while the right cylinder 5 remains sealed during its combustion cycle. Advancing an additional quarter turn to Figure 39C, the left cylinder 5 is seen blocked off during its compression cycle while the right cylinder 5 is seen connected to its exhaust port 24 during its exhaust cycle.
2o Finally, after having advanced a total of three-quarter turns, Figure 39D
illustrates the left cylinder 5 sealed off during its combustion cycle while the right cylinder 5 is seen connected to its respective intake port 21 during its intake cycle.
One of the advantages of the port distributor disk 75 is that it can be modified to allow variable port timing. Figures 40A and 40B illustrate top and bottom views of a port distributor disk 75 that consist of three concentric sections. By using control devices similar to those found in the variable valve timing systems of conventional engines, the three sections of the port distributor disk 75 can be slightly rotated in relation to one another to vary the intake and exhaust timing according to the RPMs of the engine.
Even without variable port timing, the aspiration of the Inverse Peristaltic Engine 3o is far superior to valve fed systems. However, with the engine's myriad possible applications, this feature may be deemed valuable under certain operating conditions.
Although the embodiments illustrated in figures 37-39-D use a port distributor disk 75 to control intake and exhaust, an engine using the peristaltic plate configuration that was illustrated in these figures would also function extremely well using a conventional valve setup.
Figure 41 illustrates an additional embodiment of the engine in which the connecting rods 17 are fitted with conic wheels 9 to reduce cornering wear. In the all of the vertical cylinder configurations of the engine, particularly those with small radii, the outer edges of the wheels 9 are forced to cover more distance than the inner edges that are closer to the center of the engine. Over time, this effect may eventually cause unnecessary wear to the wheels 9 and the peristaltic track 61. By allowing the wheels 9 to be conic according to the radius of the engine and by modifying the surface of the peristaltic track 61 to fit their new shape, this wear can be eliminated.
Figures 42-48B illustrate different types of open engine cooling systems. Open engine cooling systems, which are the simplest and most efficient cooling systems for use in the Inverse Peristaltic Engine, do not draw cooling air through the crankcase, so they can remain directly open to the atmosphere without discharging oil vapors.
Although Figures 42-48B only illustrate the systems as applied to the single piston per cylinder version of the engine, the single piston per cylinder version of the generator/engine unit, and the original perpendicular cylinder version of the engine, the open engine family of cooling systems may be effectively adapted to all versions of the engine and generator/engine units.
2o The open engine cooling system illustrated in Figure 42 was derived from the oil-cooled system depicted in Figures 31 and 32. Except for two disc sections, all of the portions of the ring of cylinders 5 that lie between the cylinders 5 have been removed, leaving an open area 78 surrounding the cylinders 5 (see Figure 32). In addition, two large passages 79 have been cut through opposite sides of the block. Within one of the passages 79 resides a fan 80. When the fan 80 is activated it continuously draws cool air through the opposing passage 79 and into the open area 78 surrounding the cylinders 5.
As the cool air flows around the cylinders 5, it removes excess heat from the engine before exiting as hot air through the passage 79 containing the fan 80.
Figures 43 and 44A-44C illustrate an open engine cooling system that eliminates 3o the need for an exterior fan 80. In Figures 44A and 44B the ring of cylinders 5 is seen removed from the engine. On the top disk section of the ring, four port interfaces 22 are seen positioned above their respective cylinders 5. At the center of the disk section there is seen a fifth hole 81 which allows cool air to enter the open area 78 surrounding the cylinders 5. Also seen is a number of fins 82 attached to the outer portions of and suspended between the four cylinders 5. These fins 82 may be arranged in the standard centrifugal pattern illustrated in Figures 44A and 44B or in the squirrel-cage pattern illustrated in Figure 44C. Now turning to Figure 43, there is seen a cross-section view of the engine revealing two passages 83,84 cut through the block: one passage 83 cut through the top of the block to allow cool air to enter the system and one passage 84 cut through the side of the block to allow hot air to exit. Seen again are the fins 82 that are attached to the ring of cylinders. When the engine is activated the fins 82 act as a vane and cause the air that surrounds them to rotate with the ring of cylinders. As the air's rotational velocity increases, centrifugal force creates a pressure difference between the inner and outer portions of the air's rotating mass. This pressure difference draws cool air through the hole 81 in the upper disk section of the ring of cylinders 5 and into the open area 78 surrounding the cylinders. Upon entering the open area 78, the cool air flows outward, removing heat from the cylinders 5 before exiting as hot air through the passage 84 in the side of the block. Also illustrated in Figure 43 is a baffle plate 85 positioned above the engine. The baffle plate 85 aids in the removal of excess heat from the head area of the engine by directing the incoming air flow over the top portion of the block.
Figures 45 and 46 illustrate a single piston per cylinder version generator/engine unit utilizing an open engine cooling system that cools both the engine and generator 2o portions of unit. The ring of cylinders S in this embodiment of the generator/engine unit is similar in construction to that of the engine discussed in the preceding paragraph in that all of the areas of the ring of cylinders 5 that lie between the cylinders 5 have been removed, again, leaving an open area 78 surrounding the cylinders 5. However, the ring of cylinders 5 in this generator/engine unit does not have a driveshaft 6, so a hole 81 has also been cut through the center of the bottom disk section of the ring to allow additional cooling air to be admitted to the system. Generator/engine units using open engine cooling systems may use fans 80 as illustrated in Figure 45 or fins 82 as illustrated in Figure 46 to actuate the flow of air through the system. In both cases four passages 83,86,87 have been cut through the block: one passage 83 cut through the top section of the engine portion of the unit and one passage 86 cut through the bottom section of the engine portion of the unit to allow cool air to enter the system, and two passages 87 (this number may be increased if necessary) cut through opposing lower sections of the generator portion of the unit to allow hot air to exit the system. The exit passages 87 are placed at the outer ends of the generator/engine unit so that the armature 56 in the generator portion may be cooled in addition to the cylinders 5 in the engine portion. This generator/engine unit also includes a baffle plate 85 to assist in cooling the head area of the engine portion of the unit.
Figures 47 and 48A-48B illustrate an open engine cooling system being used in the original perpendicular cylinder version of the engine. As with the previous open engine cooling systems, all of the portions of the ring of cylinders 5 that lie between the cylinders 5 have been removed to leave an open area 78 surrounding the cylinders 5. In Figure 47, several vents 88 are seen cut through the outer portion of the top section of the 1o block. These vents 88, which are also cut through the bottom section of the block, allow air to pass through the cooling system. Now turning to Figures 48A and 48B
there is seen a top and side view of a section of the ring of cylinders 5. All of the portions of the ring of cylinders 5 that lie between the cylinders 5 have been removed except for an inner ring section, a port interface 22 section and an outer ring section. Between each of the cylinders 5 there is seen a pair of fins 82. As the ring of cylinders 5 travels from left to right, the fins 82 at the top of the ring form a positive pressure above the open areas 78 surrounding the cylinders 5 while the fins 82 at the bottom ring form a negative pressure below the open areas 78 surrounding the cylinders 5. The difference in pressure causes cool outside air to flow in through the vents 88 in the top portion of the block (see Figure 47), down through the open areas 78 surrounding the cylinders 5, and out through the vents 88 in the lower portion of the block. As the cool air continuously flows over the cylinders 5, it removes excess heat from the engine.
Figure 49 illustrates an annular collector channel 89, cut into the head area of the block of the engine, which captures and re-burns any blow-by gasses that may manage to escape the interface or port area seals (port area seals apply mainly to turbocharged engines). When and if blow-by gasses enter the collector channel 89, they are immediately drawn back into the engine through a smaller channel 90 that connects the collector channel 89 to the intake area of the head. This allows the gases to be properly processed by the engine and vented through the exhaust system.
Normally aspirated Inverse Peristaltic Engines should not require a collector channel, nor should Turbo-Diesel Inverse Peristaltic Engines. (In Turbo-Diesel engines, any leakage will only be compressed air, which is not harmful to the environment.) The collector channel will be most useful in turbocharged natural gas engines or any other turbocharged engines where high manifold pressures may cause small amounts air-fuel mixture to breach the port area seals.
In Figures 50 and 51 the engine has been modified to function as an integrated centrifugal pump/engine unit. Integrated centrifugal pump units were derived from and function on the same principle as the open engine cooling system presented in Figures 43 and 44A-44C. The only difference is that the units now pump fluids instead of circulating air. To allow the pump/engine units to handle fluids, flanges 91 were extended from the top and bottom sections of the ring of cylinders 5 to provide additional room for seals. In addition, two annular collector channels 92 were cut into the block near the flange 91 to areas to capture and drain any pumping fluids that manage to breach the seals. By draining leakage from this area, the channels 92 prevent the pumping fluids from eventually seeping into the head and crankcase 15 areas of the engine. The pump/engine unit illustrated in Figure 50, which intakes pumping fluid through its top section, has retained its driveshaft 6, allowing it to function simultaneously as a stationary engine and drive other equipment as it is pumping fluids.
In Figure 51, the pump/engine unit illustrated has been fitted with a modified driveshaft 6 that functions as a directly submersible intake and eliminates the need for a sealed intake area.
Centrifugal pump/engine units do not need cooling systems because the fluid that 2o they are pumping serves as the engine's coolant. When large volumes are being pumped, the fluid will be only minimally heated by the engine. However, when it is desirable for the pumping fluid to be heated, lower volumes may be pumped and/or two or more engines may be placed in series to eliminate the need to heat the fluid elsewhere.
Figure 52 illustrates a generator/engine unit that has retained its driveshaft 6 so that it may simultaneously serve as a stationary engine and drive other equipment as it generates electricity.
Figure 53 illustrates a single piston per cylinder version of the engine with a modified bearing 64 arrangement that may be advantageous for certain applications.
Figure 54 illustrates a modified version of the oil-cooled embodiment of the 3o engine, previously depicted in Figures 31 and 32, that includes a centrifugal filler tube 93 and an air displacer tube 94 which together ensure that the reservoir area 65 remains filled with oil. In this embodiment, galleries 30 have been drilled through the driveshaft 6 to connect the filler tube 93 to the oil-filled crankcase cavity 15 and a passage 95 has been cut though the block to connect the air displacer tube 94 to the crankcase cavity 15. As the engine rotates, centrifugal force causes the filler tube 93 to draw oil from the crankcase 15 until the reservoir areas 65 are filled. To allow the reservoirs areas to 65 fill, the air displacer tube 94 provides a means for air that is displaced by the incoming oil to pass from the cooling system to the crankcase cavity 15. While less desirable, it is also possible to use a low-pressure oil pump in place of the galleries 30 and the oil filler tube 93 to maintain the proper oil level within the reservoir area 65.
Figure 55 illustrates a single piston per cylinder version of the engine in which a to number of hydraulic lifters 95 raise and lower the peristaltic track 61 to continuously vary the overall compression ratio of the engine. In this embodiment, the peristaltic track 61 has been manufactured with a number of radiating extensions 96 which slide within a number of tracks 97 cut into the inner wall of the block. The extensions 96 and the tracks 97 allow the peristaltic track 61 to be raised and lowered vertically by the hydraulic lifters 95, but prevent the peristaltic track 61 from being rotated out of position by the pistons 10. The hydraulic lifters 95 receive their commands from an electronic control unit (not shown) that takes into account a number of variables that may include the engine's workload, RPMs, temperature, manifold pressure, intake volume and exhaust gas composition in order to maximize the engine's efficiency and emission quality at any given moment.
Figure 56 illustrates a second variable compression ratio embodiment of the engine. In this embodiment, the height of the peristaltic track 61 is controlled by a number of worm drives 98 linked, synchronized, and driven by a chain 99. Power is provided to the chain 99 by one or more bi-directional electric motors 100 or by a bi-directional PTO (not shown), either of which, as with the lifters 95 in the previous embodiment, would be controlled by an electronic control unit. Seen also in this embodiment are the radiating extensions 96 and the tracks 97 cut into the inner walls of the block which work together to prevent horizontal rotation of the peristaltic track 61 while still permitting its vertical travel.
3o A third type of variable compression ratio device is illustrated in Figure 57. In this embodiment an annular screw drive mechanism is used to vary the height of the peristaltic track 61. The annular screw drive consists of a hollow, open-ended, cylinder-shaped male portion 101, inside of which the peristaltic track 61 is bolted, and a smaller ring-shaped female portion 102. The block of the engine is constructed in such a way that the female ring portion 102 of the screw is able to rotate horizontally about the engine's axis, while the male cylinder portion 101 includes extensions 96, which slide within tracks 97 cut into the inner walls of the block, to allow it and the peristaltic track 61 to be raised and lowered vertically, but prevent them from being rotated horizontally. The bottom of the male cylinder portion 101 includes protruding male screw threads that fit inside the female screw threads cut into the female ring portion 102. By rotating the female portion 102 in one direction or the other, the male portion 101 and the peristaltic track 61 can be raised or lowered accordingly. Although not illustrated in this figure, a to portion of the female ring section 102 is toothed so that it may be engaged by a hydraulically driven rack or a bi-directional, electric motor or PTO driven worm drive, or be directly geared to a PTO or electric motor. As with the two' preceding embodiments, whatever power input device is selected will be best controlled by an electronic control unit.
Figure 58 illustrates a fourth continuously variable compression ratio embodiment of the engine in which the lifting device consists of two cam rings 103,104.
The upper cam ring 103 supports the peristaltic track 61 and includes extensions 96 that slide within tracks 97 cut into the inner wall of the block to prevent horizontal rotation, while the lower cam ring 104 is constructed so that it may be rotated on the engine's axis. When 2o the lower cam ring 104 is rotated, it lifts the upper cam ring 103 and the peristaltic track 61 and hence varies the engine's compression ratio. As with the preceding figure, a portion of the lower ring 104 is toothed so that it may be engaged by a hydraulically controlled rack or by an electric motor or PTO driven worm drive.
Figure 59 illustrates a hydraulically controlled continuously variable compression ratio device that may be ideal for extreme operating conditions or simply to obviate the need for a modified peristaltic track 61. In this embodiment of the engine the peristaltic track 61 is bolted within an open-ended, hollow cylinder 105. As with the previous embodiments, the cylinder 105 includes radiating extensions 96 which slide within tracks 97 cut into inner walls of the block to allow the cylinder 105 and peristaltic track 61 to be 3o raised and lowered vertically, but prevent them from being rotated horizontally. The cylinder 105 is raised and lowered by a number of hydraulic lifters 95, as in Figure 55.
Figure 60 illustrates an opposing piston version of the engine with continuously variable compression ratio abilities. In this embodiment, the upper peristaltic track 61 has been bolted within an open ended, rotatable cylinder 106 so that it may be rotated on the engine's axis. A portion of the rotatable cylinder 106 has been toothed so it may be engaged by a hydraulically driven rack or a bi-directional electric motor or PTO driven worm drive. As the cylinder 106 and peristaltic track 61 are rotated, the timing of the upper group of pistons 10 changes in relation to the lower group of pistons 10 so that the two groups no longer arrive at their Top Dead Center positions at the same moment, thus varying the engine's compression ratio. Although this method of providing variable compression ratio abilities to the opposing piston versions of the engine is ideal, the opposing piston versions of the engine may also vary the height of one or both of their 1o peristaltic tracks 61 through the same means as one of the five preceding embodiments.
Figure 61 illustrates a fixed cylinder embodiment of the engine (see Figures 39D) that has been modified to incorporate a continuously variable compression ratio device. In this embodiment the driveshaft 6 includes a number of extensions 96 which slide within a number of tracks 97 cut into the peristaltic plate 74 to allow the height of peristaltic plate 74 to be varied while it remains engaged to the driveshaft 6. The height of the peristaltic plate 74 is regulated by a thrust bearing arrangement 107 that is coupled to a number of worm drives 98. The power needed to raise and lower the plate can be provided to the system by a bi-directional electric motor 100 or a bi-directional PTO, either of which can be linked to the worm drives 98 by chain 99. It is also possible to 2o substitute the worm drives 98 in this embodiment with hydraulic lifters 95 (see Figures 55 and 59).
Figure 62 illustrates an additional fixed cylinder embodiment of the engine which employs a simplified version of the variable compression ratio device illustrated in the previous figure. In this engine, the contacting surfaces 108 of the thrust bearing 107 and the hub portion 109 of the lower section of the block have been threaded so that the thrust bearing 107 and hub 109 may function as the nut and bolt sections of a worm drive. As the thrust bearing 107 is rotated one direction or the other, the peristaltic plate 74 and the compression ratio of the engine is raised or lowered. Although not illustrated in this figure, the outer edge of the thrust bearing 107 is toothed so that it may be engaged by a rack and be rotated hydraulically or be engaged by a worm drive and be rotated electrically by a bi-directional motor. As with all of the previous continuously variable compression ratio devices, the power input device for this variable compression ratio system would be best controlled by an electronic control unit.

Figure 63 illustrates an embodiment of the engine in which spark plugs 110 travel with the cylinders 5. Also seen in this figure are a condenser coil 111 and an electrode 112 residing within a track 113. As the spark plugs 110 travel with the cylinders 5, they act as their own distributor and are individually fired as they pass the electrode 112. The ignition timing of the engine can be retarded or advanced by varying the position of the electrode 112 within its track 113 or by eliminating the track and widening the electrode to the length of the former track 113 and signaling the coil 111 to fire at earlier or later times and for longer or shorter intervals.
Figure 64 illustrates a porous ceramic plate 114 recessed into the head of the 1o engine which can be used in place of the oil rollers or wicks 27 depicted earlier in this application (see Figure 7). The porous ceramic plate 114 is connected to an oil reservoir or the engine's lubrication system and remains saturated with oil to provide constant, even lubrication to the interface and port area seals 25,26. In addition to porous ceramic, the plate 114 may consist of a porous metal matrix or porous carbon material, or the plate may consist of a non-porous dry lubricant infused metal as was previously illustrated in Figure 9.
Figure 65 illustrates an embodiment of the engine utilizing reinforced, semi-porous metal-matrix interface and port area seals 25,26. The metal-matrix seals are connected to a low-pressure oil supply so they may absorb small amounts of oil to use for 2o self lubrication. Although not clearly illustrated in this figure, the inner surfaces of the seals are non-porous so that pressure from the combustion process will not hinder the seals' ability to absorb oil. In addition to metal-matrix, the interface seals may also be composed of fiber reinforced porous carbon or fiber reinforced porous ceramic, or may consist of a non-porous metal infused with dry lubricants.
Figures 66A and 66B illustrate an embodiment of the engine that includes stationary ceramic blades 115 within its port interfaces 22 to create swirl during the intake cycle. The ceramic blades 115 are held in place by a retainer ring 116 which is bolted to the top of the cylinder 5. Due to the fact that metal and ceramic expand at different rates when exposed to heat, an expansion gasket 117 has been placed between the retainer ring 116 and the ceramic blades 115 to prevent the ceramic insert 115 from being fractured.
Also seen in Figures 66A are the interface and port area seals 25,26. The interface seals 25 have been recessed into the retainer ring 116 so they may more closely surround the port interface. In place of the ceramic insert 115, a metal honeycomb-type regenerator 118 as illustrated in Figure 66C may be employed to reuse heat from the exhaust and aid in the vaporization of injected fuel. In this type of regenerator 118, the honeycomb pattern can be angled to also create swirl during the intake cycle.
Figure 66D illustrates a cross-section view of a recessed interface seal 25.
The interface seal 25 is held firmly to the head of the engine by a unique type of spring 119 consisting of a flat strip of spring steel that has been hardened into a crimped position.
The crimp spring 119 resists being flattened and tries to return to its crimped position and therefore provides a force to hold the seals to the head of the engine.
Figure 67 illustrates a single piston per cylinder version of the engine with a I o modified bearing arrangement that allows the driveshaft 6 to bear the entirety of the engine's rotating mass and suspend the ring of cylinders 5 so that only the interface seals 25 contact the block of the engine. (There is a slight clearance between the ring of cylinders and the block of the engine; however, the resolution of this figure is too low for it to be properly illustrated.) In Figure 68, the portlhead area of the engine has a spherical sealing surface to allow the port interfaces 22 to be moved closer to the center of the engine without compromising the engine's aspiration. By moving the port interfaces 22 closer to the center of rotation, the port interface seals' 25 average linear velocity can be halved, greatly improving the engine's sealing abilities, lowering friction and promoting longevity. In addition, the convex spherical shape of the head also provides a superior sealing surface by helping to maintain the shape of the contact areas of the seals while they are at speed. Also illustrated in this figure is a regular inner port area seal 26, a special outer port area seal 120, a special crankcase seal 121, a conventional driveshaft seal 122, and finally, a supporting ridge 123 that has been cast within the intake 21 and exhaust ports 24 to give additional support the center edges of the interface seals 25 as they pass over the intake 21 and exhaust ports 24.
Figures 69A and 69B illustrate a closer view of the outer port area seal 120 and the crankcase seal 121. The outer port area seal 120, illustrated in Figure 69A, consist of an upper female section 124, which is recessed into the head of the engine, and a lower 3o male section 125 which travels with the ring of cylinders 5. The male section 125 includes a semi-flexible protruding portion consisting of canvas, felt, leather, or other cloth-like material wrapped over or bonded to a semi-flexible steel spine. The protruding male ring section 125 fits inside of the female ring section 125 in such a way that it does not exert significant pressure on any of the female section's 124 interior surfaces unless a pressure difference between the port area and the atmosphere causes it to do so. The crankcase seal 121, illustrated in Figure 69B, which consists of a durable rubber or rubberized plastic compound molded in a "V" shape, travels with the lower disk section of the ring of cylinders 5 and seals against the provided surface on the block of the engine. The "V" shape of the crankcase seal 121 functions on a similar concept as the flexible male section 125 of the outer port area seal 120 and causes the crankcase seal's 121 sealing capabilities to increase as the crankcase 15 pressure and/or the RPMs of the engine rise.
1 o Figure 70 illustrates a peristaltic track/stationary cylinder version of the engine in which the port distributor disk 75 (see Figures 37-40B and Figures 61 and 62) has been compacted into a port distributor globe 126. The port distributor globe 126 is the stationary cylinder version's adaptation of spherical technology. The spherical shape of the port distributor globe 126 provides advantages that are virtually the same as those provided by the embodiment in Figure 68 by lowering the velocity of the areas of the globe 126 that contact the port interface seals 25 and providing a superior sealing surface.
In order to allow all of the engine's seals to be readily accessible by simply sliding the bottom portion of the engine away from the head (an Inverse Peristaltic Engine may be mounted on slide rails when installed in a vehicle so that it does not need to be 2o removed from the engine compartment to be restored), the crankcase seal 121 illustrated in Figure 71 has been relocated to the top of the lower disk section of the ring of cylinders 5 and has been provided with a flat ring 127 to seal against. In addition, an annular collector channel 128 has been cut into the block near the edge of the disk section of the ring of cylinders 5 so that it may collect and drain most of the oil slung by the lower disk section into the crankcase 15 or oil pan to improve the performance of the crankcase seal 121. Also illustrated in figure 71 is a number of traveling spark plugs 110 which may be fired using the method described in Figure 63, or by using a brush and contact ring method which will be explained in the following figures.
Figures 72A-72C illustrate two different types of brush and contact ring ignition 3o systems. In brush and contact ring ignition systems, brushes 129 extend from the ring of cylinders 5 to make continuous contact with a contact ring 130. The contact ring 130 which is attached to the head of the engine (see also Figure 71 ), directly or indirectly receives electrical power from the engine's battery or alternator. By remaining in contact with the contact ring 130, each of the brushes 129 transfers electrical power via a wire, which passes through a hole drilled through the upper disc/cone section of the ring of cylinders 5, to four individual ignition coils that reside in the open spaces between the cylinders 5. Each ignition coil then converts this electrical power into high voltage electricity to fire its respective spark plug 110. Figure 72A illustrates a top view of a ring of cylinders 5 that has been removed from the engine that was previously depicted in Figure 71. Clearly seen are four spark plugs 110, four brushes 129, four traveling electrodes 131, a stationary electrode 132 and a stationary coil 133. In the first type of brush 129 and contact ring 130 ignition system, the brushes 129 contact an uninterrupted-contact contact ring 130, illustrated in Figure 72B, that receives electricity directly from the battery or alternator. As each of the traveling electrodes 131 pass within close proximity to the stationary electrode 132, a high voltage signal spark jumps from the stationary electrode 132 to the traveling electrode 131. This electricity is then used to trip a respective relay which transfers power from its respective brush 129 to its respective traveling ignition coil, which then fires its respective spark plug 110 to begin the combustion process within a respective cylinder 5. To advance or retard the engine's ignition timing and to prolong or shorten its ignition period, the engine's electronic control unit may supply power to the stationary coil 133 at earlier or later times and for longer or shorter time intervals. Although individual ignition coils will provide the best 2o performance, the first type of brush 129 and contact ring 130 ignition system may also be adapted to use only one central traveling ignition coil instead of individual coils for each spark plug 110 by providing the central ignition coil with a direct contact to one or more brushes 129 and by placing relays between each spark plug 110 and the central ignition coil. These relays would each be tripped by signal power carried in by their respective traveling electrodes 131. The second type of brush 129 and contact ring 130 ignition system allows the traveling electrodes 131, relays, stationary electrode 132, and stationary coil 133 to be eliminated by employing a discontinuous-contact contact ring 130, which is illustrated in Figure 72C. The surface of the discontinuous-contact contact ring consists of a large neutral zone 134 that is insulated from a smaller hot zone 135.
Each of the 3o brushes 129 are directly connected to their respective ignition coil to which they can only transfer power to when they pass within the hot zone 135 of the discontinuous-contact contact ring 130. The discontinuous-contact contact ring 130 is connected to the battery by a relay, to allow the engine's electronic control unit to advance or retard the engine's ignition timing and to prolong or shorten its ignition period by supplying power to the contact ring's 130 hot zone 135 at earlier or later times and for longer or shorter time intervals. (If for some reason in the brush and contact ring type ignition systems the ring of cylinders is not able to properly ground the spark plugs through the driveshaft bearings, a grounding brush may be pressed somewhere against the driveshaft.) Figure 73 illustrates a valve-aspirated version of the peristaltic plate/stationary cylinder embodiments of the engine that were previously discussed in Figures 37-40B, 61, 62 and 71. In order to actuate the various valves 136, a cam disk 137 has been attached to the top of the driveshaft 6. The cam disk 137 should be significantly cheaper to produce than a conventional camshaft because it allows the manufacturer to simply stamp the desired cam profile onto the disk 137.
Figure 74 illustrates a valve-aspirated 136 version of the engine utilizing hydraulic lifters 95 to obtain continuously variable compression ratios (see Figures 61 and 62).
In order to reduce the length of the peristaltic plate/stationary cylinder embodiments of the engine that employ continuously variable compression ratio abilities, the outer edge of the thrust bearing 107 illustrated in Figure 75 has been threaded and placed within an annular screw drive 138 which may be rotated one direction or the other to continuously vary the engine's compression ratio (see Figure 57). In order to prevent the thrust bearing 107 from attempting to rotate, the inner portion of the thrust bearing and the hub portion 109 of the block may include splines 139, as illustrated in this figure, or the thrust bearing 107 may include one or more cylindrical extensions which slide within a comparable number of cylindrical bores drilled into the bottom section of the block.
Figure 76 illustrates another screw configuration that reduces the length of the peristaltic plate/stationary cylinder embodiments of the engine that employ continuously variable compression ratio abilities. In this configuration, which was directly adapted from the configuration illustrated in Figure 62, the inner portion of the thrust bearing 107 and the lower hub portion 109 of the block have been threaded so that the compression ratio of the engine can be continuously varied by rotating the thrust bearing 107 one direction or the other.
Figure 77 illustrates a two-cycle Diesel version of the engine. In this version of the engine a scavenging piston 140 has been attached to the lower side of each connecting rod 17. The scavenging pistons 140 reside within cylinders that are opposed to the engine's main cylinders 5 and are responsible for the intake and compression of the air that is introduced to the main cylinders 5 during the exhaust/intake cycle via scavenging ports 141. (The scavenging piston's intake and compression cycles are managed by reed or spring type check valves which are not illustrated in this figure.) By making the bore of the scavenging cylinders larger than that of the main cylinders 5, the engine can be supercharged by advancing the closing of the exhaust valves 136. In addition to eliminating the need for a scavenging pump, the opposed-cylinder/two piston per connecting rod design of this engine also imparts additional stability to the engine by eliminating the prying force exerted on the connecting rods 17, pistons 10 and cylinders l0 5. If the scavenging ports 141 are eliminated, and intake valves are added to the upper portion of this engine so that it may operate as a four-stroke engine, the former scavenging piston 140 and check valve system may be modified to function as an integrated gas compressor.
Figure 78 illustrates another two-cycle version of the engine that uses conventional scavenging methods.
Figure 79 illustrates an opposed cylinder configuration that was derived from the two-cycle version of the engine that was previously depicted in Figure 77. In short, this version is simply two engines sharing the same connecting rods 17, peristaltic plate 74 and driveshaft 6. While this configuration, which can also easily be adapted to the 2o traveling cylinder versions of the engine, does place some restrictions on the flexibility of the peristaltic plate 74, it may prove valuable were there is a need for a narrow, high powered engine. As with the two-cycle opposed-cylinder version discussed earlier, this opposed-cylinder arrangement also imparts additional stability to the engine by eliminating the prying force exerted on the connecting rods 17, pistons 10 and cylinders 2s 5.
Figure 80 combines the time-tested sealing ability of poppet valves 136 with the superior aspiration of a port distributor globe 126. In this engine each cylinder has one large poppet valve which is actuated by a cam disk 137 that extends from the port distributor globe. The poppet valve 136 is only responsible for sealing during the 3o compression and combustion cycles. The intake and exhaust cycles are modulated by the port distributor globe 126. Poppet valves 136 can also easily be used in unison with the port distributor disk 75 that was illustrated previously in Figures 37-408, 61 and 62.
Finally, in Figures 81A-81D an isolated traveling cylinder 5 is depicted on a peristaltic track 61 at different consecutive periods in time to illustrate the versatility of the peristaltic track 61 and to further clarify the peristaltic process. In Figure 81 A the traveling cylinder 5 is depicted during its intake cycle and is currently drawing air/fuel mixture in to the cylinder 5 through the intake port 21. In Figure 81 B the traveling cylinder 5 is depicted during its compression cycle. In Figure 81 C, the spark plug 110 has fired and combustion is currently underway within the cylinder 5. In this particular peristaltic track 61 pattern, the stroke length of the combustion cycle is approximately twenty percent longer than that of the intake and compression cycles, allowing prolonged/extended expansion, and therefore greatly improving the engine's overall 1o efficiency and power output. In Figure 81D, the cylinder 5 is depicted during its exhaust cycle and is presently expelling its exhaust gasses through the exhaust port 24. Upon completion of the exhaust cycle, the cylinder 5 will begin to pass through each of the four cycles again.
ADDITIONAL COMMENTS
Although electronic control units (computers) will provide the best results, all of the continuously variable compression ratio devices described in this patent application may also be controlled mechanically.
The splines or "radiating extensions 96 and tracks 97" used within all of the continuously variable compression ratio embodiments of the engine, may be slightly 2o twisted, angled or corkscrewed so that the port or valve timing of the engine may be varied with the compression ratio.
The generator/engine units in this application may also include continuously variable compression ratio abilities by employing the same devices used in the continuously variable compression ratio embodiments of the engine.
Some of the engines in this patent application were depicted with two fuel injectors at each repeat in the peristaltic pattern. This number may be increased or decreased depending on the injection time needed at the engine's ideal operating speeds.
To extend the life of the block of the engine and the ring of cylinders and to expedite and improve the quality of overhauls, a replaceable hardened steel liner may be 3o bolted onto the peristaltic track and any other wear prone areas of the engine to eliminate the need to regrind the surfaces of the engine's contacting moving parts.
The same Inverse Peristaltic Engine can be fueled by gasoline, Diesel, recovered lubricating oil, propane, crude oil, natural gas, methanol, ethanol, Bio-Diesel, landfill gas (methane), hydrogen or any other oily, liquid or gaseous fuel.
The Inverse Peristaltic Engine can be geared to the desired operating speed within the engine by modifying the slope of the contoured surfaces of the peristaltic track, thus eliminating the need for an external transmission.
EFFICIENCY AND FLEXIBILITY
The Inverse Peristaltic Engine has no true Top Dead Center or Bottom Dead Center. Transitions from dead points to areas of maximum mechanical advantage are almost instantaneous. This allows the engine to waste very little thermal energy because it can convert the majority of its thermal energy into mechanical energy at a time of maximum cylinder pressure.
The unique configuration of the Inverse Peristaltic Engine also allows the expansion or compression rate, stroke length, time duration, and compression ratio of each of its four cycles to be fully adjustable and to vary from one cycle to another. In addition, the overall compression ratio of the engine can be continuously adjusted as the engine is operating to maximize its efficiency at any given workload or RPMs.
Combined, these elements and those in the paragraph above make it possible to capitalize on almost all current knowledge regarding engine efficiency.
ASPIRATION
The port system used within the Inverse Peristaltic Engine will provide a level of 2o aspiration which is far superior to that of any valve aspirated engine. In most cases, a naturally aspirated Inverse Peristaltic Engine should intake more air than a supercharged conventional engine. In addition to providing a high horsepower per liter ratio, the port system will also reduce pumping losses as well as losses to engine vacuum due to delayed intake timing or simply the peak profile of a camshaft (compression is not a true loss).
PROLONGED EXPANSION
In a typical crankshaft engine, especially those that are turbocharged, a significant level of pressure remains in the cylinder at Bottom Dead Center only to be wasted in the exhaust. However, in the Inverse Peristaltic Engine, the length of the combustion stroke can be made longer than the intake and compression strokes in order to make full use of 3o this residual energy. In addition to initial power gains, the lower pressure left in the cylinders will allow the engine to use less energy to expel its exhaust. The lower pressure during the exhaust cycle will also allow more exhaust to be expelled, thus allowing more air to be accepted during the intake cycle, further increasing the performance of the engine.
HORSEPOWER DENSITY
The Inverse Peristaltic Engine's extremely high efficiency combined with its superior aspiration should allow the engine to reach enormously high power densities. A
naturally aspirated Inverse Peristaltic Engine should breath as well as a supercharged conventional engine.
TIMMING
The Inverse Peristaltic Engine is timed by adjusting either the number of four-stroke peristaltic pattern repeats or the number of cylinders, or both, until the multiples of to the number of peristaltic pattern repeats divided by 360 degrees do not equal any of the multiples of the number of cylinders divided by 360 degrees until the multiples reach 360.
MECHANICAL ADVANTAGE COEFFICIENT (TORQUE) AND RPMs In a crankshaft engine, the crankshaft must travel two rotations to complete four strokes; therefore, if the stroke lengths of both engines are equal (in versions of Inverse Peristaltic Engine that contain opposing pistons, the stroke of each piston must be added together to yield the net stroke), the RPMs of the crankshaft engine divided by twice the number of the Inverse Peristaltic Engine 's four stroke peristaltic pattern repeats equals the RPMs of the Inverse Peristaltic Engine. This may be transposed to indicate the mechanical advantage coefficient of the Inverse Peristaltic Engine. Again, if the stroke lengths of each engine are the same and if the slope of the Inverse Peristaltic Engine's contoured peristaltic walls is the same for each of its four cycles, the number of times more torque that an Inverse Peristaltic Engine has than a crankshaft engine equals twice the number of four stroke peristaltic pattern repeats. The opposing piston version engine depicted in Figure 1 has twelve pattern repeats, so the mechanical advantage that its pistons have over its driveshaft is twenty-four times that of an equivalent crankshaft engine, and its RPMs are twenty-four times slower. For the most part, the net slope of the Inverse Peristaltic Engine's peristaltic track is irrelevant. If stroke length and number pattern repeats are held stationary, varying the net slope of the peristaltic track only affects the radius of the engine. While increasing the slope increases mechanical advantage at the track, it causes an equivalent reduction in the engine's lever arm, which prevents the engine's torque from being changed. This is explained by the following equation:

Power slope P - Power stroke Intake slope ~ - Intake stroke C Compression slope c - Compression - stroke E Exhaust slope a - Exhaust stroke -to R Radius of track ~ - Sum of X,Y,Z....
-Radius of crankshaft S - Stroke ( 2 r = s ) l5 (2~R)_((I/taneofi)+(c/taneofC)+ (e/taneofE).
20 or P/taneofP
25 I radius C'rankchaft Peristaltic Track (One pattern repeat) R-(~(i/taneofi),(c/taneofC),(P/taneofP),(e/tanBofE))/(2x~) Since tan a falls in the denominator, an increase in a causes a equal and opposite decrease in R.

SUBSTITUTE SHEET (RULE 26) A more simplified proof can be arrived at by keeping the length of track length of each of the four cycles equal:
.adius / 2 (~rank~haft radius Peristaltic Track (One pattern repeat) ( 2r l tan 8 ) ( Pattern Repeats ) _ ( T~ R ) l 2 Which simplifies to:
((4r)x(Repeats))/((tane)x(~)) Again, tan 8 falls in the denominator, so any increase in a causes an equal and opposite decrease in R. While increasing a increases mechanical advantage at the track, the reduction in R, or the engine's lever arm, prevents the engine's torque from being changed.
3o For the versions of the Inverse Peristaltic Engine containing opposing pistons, the previous equations match the cycle speeds of each engine, but not the piston speeds. In cylinders with opposing pistons, each piston travels only one-half of the total stroke and therefore only one-half of the cycle speed; so, for an Inverse Peristaltic Engine using opposing pistons, the RPMs of the crankshaft engine must only be divided by one times the number of four stroke peristaltic pattern repeats for the RPMs of an Inverse Peristaltic Engine whose piston speeds match that of the crankshaft engine.
This equation, which will double the RPMs of the previous equations, should help predict the maximum tolerated RPMs of the opposing piston versions of the Inverse Peristaltic Engine. However, because the cycle speeds are not equal, this equation cannot be used to 4o derive a mechanical advantage coefficient.
From the equations above, it is clear that changing the number of peristaltic pattern repeats will vary the RPMs and torque output of the Inverse Peristaltic Engine.
However, these parameters can also be adjusted by modifying the slope of the peristaltic track that the cylinders pass over during the combustion cycle. In order to realize a change in mechanical advantage, modifications made to the combustion area must be made by changing its slope in relation to the slope of at least one of the areas passed over during the three other cycles. In this case, the following equation must be used to determine the engine's mechanical advantage coefficient in relation to a conventional engine with equal stroke:
to (( tan a of power slope) x ( radius of track )) / (( stroke / ~ x radius of crankshaft ) x ( radius of crankshaft )) Which simplifies to:
(( tan 8 of power slope) x ( radius of track )) / (( 2 x radius of crankshaft) / ~)) The radius of the track being determined by:
[ (( intake stroke / tan 8 intake slope ) + ( compression stroke / tan 8 of compression slope) + ( power stroke / tan a of power slope) + ( exhaust stroke / tan a of exhaust 2o slope)) x ( pattern repeats ) ] / ( 2 x ~ ) If as a result of this modification the walls that the cylinders pass over during the compression cycle differ in slope from the walls that the cylinders pass over during the combustion cycle, an engine constructed in this manner running clockwise will demonstrate RPMs and torque outputs that are inversely proportional to those experienced when the same engine is run counterclockwise; therefore the manufacturer only needs to build one type of engine to fulfill the needs of two different applications.
While providing a true mechanical advantage coefficient, the equations that have been listed in the paragraphs above are only useful to provide an estimate of the torque outputs of the Inverse Peristaltic Engine when compared to that of conventional types.
The complete torque equation (neglecting friction and other associated losses) for one full rotation of a four stroke Inverse Peristaltic Engine, regardless of its number of peristaltic repeats, is:
[(Average force input at piston) x (tan g of power slope) x ( # of cylinders) x (radius of track)] / 4 If the slope and/or stroke of the track is different for different cycles, the equation is:
[(Average force input at piston) x (tan 8 of power slope) x ( # of cylinders) x (radius of track)] x [( power stroke / tan a of power slope) / (( intake stroke / tan g intake slope) + ( compression stroke / tan a of compression slope) + ( power stroke /
tan a of 1o power slope) + ( exhaust stroke / tan g of exhaust slope))]
EMISSIONS
The Inverse Peristaltic Engine is compatible with all preexisting emission technologies; however, its unique configuration gives it countless advantages over conventional engines. The Inverse Peristaltic Engine's near instantaneous transition from Top Dead Center reduces stall time during the combustion cycle and allows the engine to rapidly convert its thermal energy into mechanical energy. This rapid energy conversion reduces the length of time that the combustion cycle remains at peak temperature and thus greatly reduces the formation of NOx emissions. The peristaltic configuration of the Inverse Peristaltic Engine also allows the expansion or compression rate, stroke length, 2o time duration, and compression ratio of each of its four cycles to be fully adjustable and to vary from one cycle to the next. Additionally, the overall compression ratio of the engine can be continuously adjusted as the engine is operating to produce clean emissions at any given workload or RPMs. These features make it possible to capitalize on almost all current knowledge regarding engine emissions.
Without question, the peristaltic configuration of the Inverse Peristaltic Engine will be capable of yielding the highest relative torque output ever produced by an internal combustion engine. However, the defining traits of the Inverse Peristaltic Engine are its unique method of power conversion and its unprecedented flexibility of design.
The engine's method of power conversion allows it to achieve stratospheric efficiencies while its incredible flexibility of design grants the manufacturer complete, uncompromising control of the engine's torque, emission levels, RPMs, and the individual behavior and nature of each of its four cycles. It is my hope as the inventor that this patent will pioneer a new frontier in the world of engine technology that will lead to cleaner and more efficient methods of harnessing the energy that makes the world go round.
The foregoing embodiments are presented by way of example only; the scope of the present invention s to be limited only by the following claims.

Claims (13)

CLAIMS:

1. An inverse Peristaltic Engine apparatus comprising:
a) a plurality of interconnected cylinders containing pistons engaged to a peristaltic member;
b) means for admitting and expelling fluids from the cylinders; and c) means for igniting the contents of the cylinders to power the engine.

2. The apparatus in claim 1, wherein the means for admitting and expelling fluids from the cylinders comprises valves.

3. The apparatus in claim 1, wherein the means for admitting and expelling fluids from the cylinders comprises ports.

4. The apparatus in claim 2, wherein the valves are actuated by means of a rotating disk or wheel.

5. The apparatus in claim 3, wherein the ports are sealed against a spherical head area of the engine.

6. The apparatus in claim 5, wherein the spherical head area further comprises secondary seals and channels to control emissions.

7. The apparatus in claim 1, wherein the engine has an overall compression ratio that may be continuously varied.

8. The apparatus in claim 1, wherein there are further provided collector channels to control the leakage of oil.

9. The apparatus in claim 1, wherein engine combustion is controlled by a brush and contact ring ignition system.

10. The apparatus in claim 1, wherein the peristaltic member further comprises a plate member.

11. An inverse peristaltic engine, comprising:
a) a plurality of interconnected cylinders containing pistons engaged to a peristaltic track;
b) valuing means for admitting and expelling fluids from the cylinders; and c) means for igniting the contents of the cylinders to power the engine.

12. The apparatus in claim 11, wherein the means for admitting and expelling fluids from the cylinders comprises ports.

13. An inverse peristaltic engine having a continuously varying overall compression ratio, the engine comprising:

a) a plurality of nterconnected cylinders containing pistons engaged to a peristaltic plate;
b) port means for admitting and expelling fluids from the cylinders; and c) ignition means for igniting the contents of the cylinders to power the engine.