ROTARY DRIVE MECHANISM

Abstract: A rotary drive mechanism for use as a motor, a pump or a compressor includes a housing (13) having a chamber (1) therein defined by a peripheral wall. A rotor (3) is rotatably mounted in the chamber (1) and has two longitudinal seal edges (5) that contact the profiled wall. The rotor (3) is mounted for rotation about a rotation axis (9) and for sliding movement relative to the rotation axis in a direction perpendicular thereto. The rotation axis (9) is offset form the centre of the chamber.

Description: ROTARY DRIVE MECHANISM The present invention relates to a rotary drive mechanism and a rotary displacement geometry for use as a motor, pump or compressor. Another aspect of the invention relates to a variable delivery pump mechanism and a further aspect of the mvention relates to a heat cell. According to one aspect of the invention there is provided a rotary displacement geometry for use as a motor, pump or compressor, in which the simple motion of a two sided rotor sliding through a displaced axial centre creates compression ratios of between two and fifteen to one, at maximum displacement, the chamber profile being a loop which could be circular, part circular or fully non-circular, in each case the profile will be contacted by either the two longitudinal rotor edges or seals mounted there upon, the axial displaced centre being offset from the actual mid-centre by less than one-sixth of the chambers effective internal diameter, and the rotor's circumscribed foil cross sectional area being 30% or more of the chambers full cross sectional area. According to a further aspect of the invention there is provided a rotary drive mechanism for use as a motor, a pump or a compressor, the mechanism including a housing having a chamber therein defined by a peripheral wall, a rotor rotatably mounted in the chamber having two longitudinal seal edges that contact the profiled wall, the rotor being mounted for rotation about a rotation axis and for sliding movement relative to the rotation axis in a direction perpendicular thereto, the rotation axis being offset from the actual midway centre of the chamber. _ , According to a further aspect of the mvention there is provided a variable delivery pump mechanism including a housing having a chamber therein defined by a peripheral wall, a rotatable rotor mounted in the chamber and dividing said chamber into a plurality of sub- chambers, the rotor being constructed and arranged such that rotation thereof causes the volume of said sub-chambers to increase and decrease alternately, and an inlet/outlet port member having at least one inlet port and at least one outlet port therein, said inlet/outlet port member being rotatably adjustable relative to the housing, to adjust the delivery volume of the pump. Within the text of the application, there are three main novel devices with applications in their own right. When combined, they can avail themselves as a means of mechanical propulsion for powering vehicles etc. The devices are a means of heating free or pre- compressed air to power rotary heat engines, and rotary motor/pump design geometry and a means of varying pump delivery by retarding ports rotationally. Well-insulated heat cells are a means of storing energy, mainly heat charged from electricity for the purpose of powering heat engines. The rotary motor/pump is a novel but simple geometric approach that seals and moves in an efficient manner. The rotary geometry has applications in fluid or gas as a pump, compressor or motor as well as in heat or internal combustion engines. Although there are numerous references to heat engines and heat cells, it must be appreciated that mostly the key to all, is the novel rotor geometry. The rotor geometry overcomes the main flaws of existing rotary design; the two main ones being excessive oil requirement and rotor wear on the seals. It will be seen that the proposed rotary device allows sealing characteristics equivalent to piston rings, whilst offering all the benefits of rotary motion. A large emphasis has been put on the applications in motor vehicles, primarily because of environmental concerns. With heat engine systems, the preferred method is the induction and heat expansion of an air/gas mix, much the same as in power generating gas turbines which will normally be closer to 50% efficient. There are two main rotary geometries which operate within a chamber, the simplest form is an arrangement of one or more radial projecting vanes which slide in and out or through a slotted hub on a displaced centre within a circular or non-circular chamber. The vane movement is considerable and so too are the lubrication requirements, in terms of air compressor applications; lubrication becomes a nuisance albeit not insurmountable. The proposed rotary geometry on the other hand has its lubrication and sliding components internally enclosed within the rotor itself. Another form of rotary displacement pump/motor uses a rotor with two, three or even four active rotor sides; the rotor orbits and gyrates upon a reduced diameter toothed shaft. The best-known example of this rotary geometry is an internal combustion engine called ' the Wankel engine. This engine operates by a triangular rotor gyrating around a figure of eight chamber periphery. The rotor meshes onto a hollow externally toothed shaft with the toothed bore of the rotor being a significantly greater diameter than the shaft. The toothed gearing thrusts the rotor to hoop around the shaft in a gyrating cycle within the chamber geometry. The movement causes live volume displacement between any two of the rotor tip faces on each of the three rotor sides. This creates the four cycles of internal combustion at set arcs around the chamber. The motion of the rotor causes undue wear on the rotor tips causing a limited life, however, even with this drawback and associated fuel port bypass problems, the engine creates enormous compact power. The proposed rotary geometry has a chamber profile that can provide smooth acceleration of rotor motion. The motion can be manipulated to deliver optimum rotor lever arm characteristics dedicated to extract maximum power efficiently. Such Wankel motion is overly excessive and causes acute tangential angle change between the seals and the chamber wall inevitably causing seal wear. At the turn of the last century, there were a number of pump/engine patents that incorporated a similar chamber profile, though the rotary component was fundamentally flawed. The rotary component consisted of a single sliding vane that slides through a rotating boss or shaft. Such arrangements by their nature, had no facility to outwardly radially seal, and required inoperable amounts of lubricant exposed to the chamber to lubricate the sliding action. The proposed arrangement overcomes all of the above- mentioned flaws and retains a smooth motion through a stable axis. The proposed two-sided rotor has two floating edge seals and achieves its volume displacement by sliding through a displaced centre as it rotates. The radial enlargement of the rotor's peripheral cross-section taking up 40% or more of the chambers full internal cross sectional area helps to enhance displacement, enabling the displaced centre off set to be less than l/6th the chamber diameter. The rotor actually slides through the displaced centre in a restrained manner, reducing the need for seal movement. The chamber will preferably be of a dedicated geometry so as to maintain a virtual zero gap with the rotor. This would provide a rotaiy device, say in ceramics, theoretically avoiding the need for seals. However, even if the rotor does not follow such a dedicated peripheral geometry, the rotor itself will accommodate the major sliding travel of up to l/3rd the effective chamber diameter with seals only traversing up to l/20th of diameter distance. Such rotor bearing restraint allows the small sliding travel to be smooth and controlled in a manner of uniform acceleration and deceleration. The sliding movement takes place through a slotted boss that is mounted in the main bearing at either end, axially upon the displaced centre. The rotor movement is constrained by bearings acting on a inward set track corresponding chamber peripheral, the edge seals are effectively floating with exceptional wear life potential. The vast bulk of mechanical contact and loading is carried out through roller bearings providing exceptional wear life and a minimum of mo ving parts . The rotor in the proposed arrangement contains all its lubricant requirements internally. The reduced motion and travel is achieved by enhanced rotor enlargement, this means that multiple bank edge seals can be utilised without undue fluttering on acute chamber tangent angles. Additionally, the edge seals if singular, can be of significant section thickness allowing them to act as the impelling bearing surfaces that contact on the chamber wall. This in turn allows the motion to be governed around a simple scotch crank. The proposed rotary motor/pump has a highly unique overall geometry, capable of delivering compression ratios of 2:1 up to 15:1. In a more elaborate version the motion is governed via a bearing set at either end within the end caps. This not only avoids high pressure wear contact on the tips of the rotor, but also causes the seals to act efficiently in that they are then isolated from the rotors dynamic loading. The peripheral geometry of the chamber may be circular, part circular or fully non circular. In the case of the latter, the peripheral geometry is of dual ellipse where the lower portion of the chamber is an ellipse split along the long axis. The upper chamber periphery is derived from and creates another half ellipse tracked from the lower when traversed through the displaced centre at a rotor's tip-to-tip distance. This geometry is ideal for driven devices such as pumps and compressors as it provides the optimum smooth rate of rotary motion. Variable delivery arrangements such as an oil pump provide means of adjusting and controlling delivery to match engine requirements, thus improving the overall energy efficiency in engines. Engines require a significantly (up to 60%) lower linear output of oil delivery per engine rotation cycle at high speed than they do at low speed. As engine speed increases therefore the oil delivery rate per engine rotation cycle must be reduced proportionately in order to balance oil flow rates to specific engine requirements. There is a distinct imbalance between the performance of conventional gear profiled oil pumps and precise engine needs, especially at higher speed, when pump output increases and delivery requirements per engine cycle diminish. Accordingly excessive oil flow results at high engine speed, this excess is usually released into the sump causing significant turbulence, mist and foaming within the lower engine sector. Such excessive delivery output causes a situation by which these conventional oil pump arrangements can absorb up to as much as 4% of the engine's total power output. The preferred arrangement comprises the use of a two sided rotor that will not only act in a more positive manner but will also provide a degree of precise control vis-a-vis rotational port manipulation. There are many other pump and compressor applications where variable delivery provides benefit, for example with air compressor where the electrical drive motors do not run efficiently in a stop start manner. Another example is a refrigeration compressor on the air conditioning unit of a car, where the engine speed varies and the internal car cabin temperature itself is a variance. Within the context of the invention as proposed there is provided a means of resolving the problem of the inefficiencies of excess delivery by incorporating port manipulation. There is a means by which oil pump flow rates can be manipulated via a distortion or rotation of the port settings to advance or retard the point upon which induction and compression is enacted and thereby vary the effective displacement and delivery volume as required. The preferred port manipulation takes place upon both ports. By introducing one or more plates/screens, incorporating ports, that can be varied rotationally to alter the point upon which displacement is enacted, the priming volume can be manipulated to back feed on itself with negligible resistance. Preferably the rotational slide arc of such a port plate will be around 80° and will allow an output variance of around 2.5 to 1 of delivery on a standard rotor displacement. Similar variable delivery mechanisms, in the form of adjustable plates or screens can be provided to this and other types of rotary displacement pumps, which may be multi lobed and chambered such as in the form of a rotary gear pump. In the case of an engine/motor being a driven device the radial lever arm will need to be disposed toward the extended volume portion of the chamber to accentuate and absorb the power input. In this case the lower periphery will be somewhat parabolic to accelerate the rotor's radial lever arm into the upper chamber. The profile may in fact be part or wholly non symmetrical about the displaced centre, or in part follow a circular geometry for 130 degrees of angular rotation. This allows the extended radial lever arm to arc for a good period of travel. In consideration of engine uses, the form could be used as an internal combustion engine with two or more chambers banked together, with say the first pressurising and super- charging the other with fuel air mix. The two units could also be combined to run in tandem, backed off against each other, to create a balanced cycle. A layout could be applied to heat engine units with a right angle rotor alignment between the two-banked units. The right angle alignment would provide a receptive rotor inclination allowing instant pressure start at any rotational rest angle, as well as balanced power output. A preferred version of the rotary arrangement as a heat engine in the format of a double expansion chamber powered by pre-compressed hot gas, much in the same way as a gas turbine engine. The air intake would be supplemented to increase flow and operating pressure by either rotary pump or turbo fan. The benefit of such an arrangement is that as with gas turbines there is a far greater power output related weight and engine volume. The advantage in efficiency in comparison to gas turbines is that the output of gas pressure is contained and absorbed through a fully sealed chamber. This allows the pressures to be fully absorbed in a controlled and precise manner at significantly lower engine speeds than turbines. The proposed arrangement would allow substantially more heat/pressure energy to be absorbed and utilised exhausting far lower temperatures. This arrangement also has the inherent advantage over conventional internal combustion engines in that the power throughput is of a constant pressure flow rather than less efficient intermittent combustion firings. Conventional internal combustion engines, though balanced, lose a lot of energy associated with friction and the numerous moving parts involved. Conventional internal combustion engines also have the limitation of few specific fuel types that are chemically suitable for compressive combustion. The proposed arrangement would provide comparable if not potentially much greater efficiency than conventional engines. The cleanest fuel would obviously be hydrogen and oxygen/air; the combusting characteristics not being compatible with internal combustion would however suit the proposed arrangement. The proposed arrangement would have an air heater arrangement that transfers the heat from a remote main source into the pressurised airflow to deliver significant pressure. By pre-compressing the heated gas to say two bar the heating effect can double the working pressure up to around four bar at temperatures of just over 520°C. This provides an active working pressure three Bar at average global atmospheric temperatures. Even at these relatively low working temperatures /pressures this would provide an engine unit of a lesser volume and weight than conventional engines. If required the main components could be produced in ceramics to operate more reliably at significantly higher temperatures. An example of the extent of power that can be derived from similar low-pressure device is the considerable output obtainable from small compressed air driven motors. The proposed arrangement lends itself to all potential fuels (those that deliver heat energy). There is strong argument that this is a more practical, versatile and powerful approach to that of alternative fuel cell technology. In terms of logistics the proposed arrangement has the ability to be in a dual fuel format making it a viable proposition for existing fuelling station infrastructure The main drawbacks with existing vehicles are basically that internal combustion engines are only around 20% efficient and electric battery vehicles are excessively heavy. Electricity is clean, and the energy prices, at off-peak are around one seventh that of petrol. Potentially, however, there are more immediate prospects of increasing efficiency of conventional vehicles by switching to external combustion. Internal combustion engines have to expel large amounts of heat energy, whereas external combustion engines actually utilise this heat. By combining this approach with a heat cell the main fuel could be pre- charged electrical heat, there would be environmental benefits. The weight of a heat cell to achieve a ninety-mile range at fourteen horsepower (lOkw) would be only 200kg. The extra weight would be inconsequential in comparison to the enormous foel cost and environmental benefits. Electric battery vehicles have over 500kg of battery for half this range. Heat cell power to weight ratios would provide performance characteristics comparable to petrol vehicles. Such vehicles could be pre-charged overnight with pre- required energy levels to further maximise the efficiency of the following day's travel. It would be possible to achieve a match in conventional vehicle weights and still retain a 100kg heat cell by virtue of the reduced engine plant weight. This would create a dual fuel vehicle capable of short runs of 40 miles plus, on electric heat driven and the longer runs by external combustion fuelling. Related to the viability of the heat cell powered vehicle detailed calculations reveal the following: Conventional re-chargeable vehicle batteries only hold 0.01765kw hours per kg. Magnetite heat cells at 800°C hold 0.149kw hours per kg, this represents an 8.4 benefit by weight. A 200kg heat cell would provide a vehicle range of 90 miles at 14 horsepower (1 Okw). Heat engine performance characteristics would be comparable to petrol vehicles. Heat cells have an indefinite life expectancy when compared with vehicle batteries. Fuel running costs using off-peak electricity would be l/7th that of current UK petrol pricing. A 90 mile heat cell (200kg) would measure 450mm x 450mm x 450mm, the size of a portable television (16" x 16" x 16") including the casing and insulation. The heat energy retained would be 92.7% over 18 hours and 71% over 72 hours. This assumes the worse energy loss (efficiency) case scenario of a folly charged heat cell where there is no positive energy extraction within the period. Embodiments of the invention will now be described by way of example with reference to the accompanying drawings, in which: Figures 1 and la show two alternative chamber profiles (1) and roller track (11). Figure 2 shows the construction method of one type of chamber profile (1) and one rotor roller (7) in contact with the bearing track (11). Figure 3 shows the rotor (3) assembly with the slide bars (8) and rollers (7) connected, and slide boss (10) to the right. Figure 4 shows the rotor (3) assembly with the bearing assembly in place and the possible extent of compression displacement. Figure 4a shows the more basic rotor (3) assembly that rotates on a shaft (68), the motion of the rotor (3) being that of a scotch crank. Figures 5, 6, 7 and 8 8A, 8B 8C and 8D show the edge seal (5) alternatives and contact angles in comparison to proven piston rings (22). Figure 9 shows the end cap and bearing details. Figures 10, 10A and 10B show the action of the rotor (3) driven under pressure. Figures 11, 11A, 11B and 11C shows a simple scotch crank slide rotor (3) pump/compressor running through three stages of its displacement cycle. Figures 12, 12C and 12C shows the effect of rotationally retarding the ports (43,44) on volume delivery/real displacement and the flow effect on a double-banked rotor (3) arrangement. Figure 13 shows a dual rotor chamber Figures 13a, 13b, 13c, 13d and 13e show a dual rotor (3) vapour/steam cycle Figure 14 shows a dual rotor (3) arrangement for use as a double expansion hot gas or vapour engine and port (43,44) arrangements. Figures 15, ISa and 15b 15c 15d and 15e show a double expansion rotor (3) cycle. Figures 16, 17 18a and 18b shows various heat cell (55) arrangements and ancillaries. Figures 19a, 19b and 19c show a double expansion hot gas rotor (3) and arrangement in isometric view. Figure 20 shows a comparison graph between heat engines and internal combustion. Figure 21 shows a comparison graph comparing heat cells to rechargeable vehicle batteries. Figure 22 shows an alterative form of rotor pump. Figure 1 shows the unique chamber profile ( 1 ) and the inwardly reduced bearing track (11) enabling the rotor (3) shown in Figure 3 to rotate within the displacement chamber, the bearing track (11) setting the contact distance between the rotor (3) tips and the chamber (1) with a parabolic configuration. The midway centre (33) emphasises the extent of displacement upon the active centre immediately below. Figure 1 A shows another chamber (1) profile format primarily used as a pump/compressor similar to that in Figure 1. Figure 2 shows how the unique chamber profile (1 ) provides an accurate and equal distance (2) at any point of rotation. The unique chamber geometry takes the form of an exact ellipse below the centre line construction around the locus points (50), the upper profile above the centre line is an ellipse defined by tracking chamber width distance (2) through the displaced centre. The roller bearing (7) can be seen at a set distance from the chamber profile (1) running on an internal peripheral track (11). This figure also demonstrates how the roller track (11) could be machined by a simple jig carrying a milling tool of exact corresponding roller (7) diameter, held at an exact centre distance to machine profile of reduced proportion to the chamber (1) to create a rotor (3) roller (7) control gap with the chamber (1). This allows the rotor (3) to rotate without coming into contact with the chamber (1). Figure 3 shows the components of the rotor (3) assembly comprising of the rotor (3) the motion governing roller bearing (7) the slide bar (8) and the slide boss (10) that together permit sliding travel through the displaced centre, whilst restraining gyrational loadings. Figure 4 shows the rotor assembly engaged via rollers (7) within roller track (11) with the slide boss (10) in its axially displaced centre/position (9) engaged over the slide bar (8) with seals (5) contacting the chamber (1), whilst retaining a gap with the rotor (3) tips. Figure 4A shows the basic form of rotary device, where the rotor (3) simply slides on a flat shaft (68) which is set into the main bearing (18) at either end on the displaced centre (9). The flat shaft (68) is housed in a void within the rotor (3). The edge seals (5) can be seen to be in contact with the end seal (16) that applies an outward pressure to the edge seals (5). Figure 5 shows a conventional piston (21) cylinder (20) arrangement with the effective sealing action of piston rings (22). It can be seen and appreciated elsewhere that the rotor edge seals (5) act in an equally efficient manner, whilst retaining all the benefits of rotary non-reciprocating motion. 5 Figure 6 shows the application of a single seal (5) which other than slight spring pressure is effectively floating much in the same way as the piston ring (22) of Figure 5, and where a small gap is maintained between the rotor (3) and the chamber (1). The edge seals (5) benefit from self-mass centrifugal forces. Figure 7 shows the application of twin seals (5) cradled in a spherical rocker (6) to reduce 0 seal (5) movement further if required. This would achieve adequate sealing between the rotor (3) and the chamber (6) with nominal seal (5) movement where the spherical rocker ■ (6) floats on a film of oil pressured by centrifugal force. Figure 8 shows the arrangement by which three seals (5) are set in seal bushes (67) allowing them freely to contact the chamber (1) under the light outward pressure of the end seal (16) 5 movement within the slackened seal recess (69) which is lubricated by the rotor (3) having a small oil port (34) feeding the edge and end seals (5) and (16). Figure 8A shows the tangential angle (48) effect of edge seal (5) surface contact upon the chamber wall (1). The small tangential angle (48) change is shown against the optimum right angle (47). The line upon which the right angle line (47) is draw from the rotor (3) 0 axis line (42) at four points around the chamber (1) on the right hand side. It will be seen that the uppermost tangent -line (48) sits at the optimum zero to the perpendicular line (47) and the same occurs opposite bottom end. Figure 8B shows a single edge seal (5) arrangement bushed (49) both sides and the capacity • of the arrangements to accommodate the tangential angle change (48) through radiusing 5 edge seals (5). Figure 8C shows a dual edge seal (5) arrangement set in bushes (49) and the capacity to accommodate the marginal fluctuating travel distance (51). Figure 8D shows a similar arrangement to Figure 8C with a triple bank of edge seals (5), the fluctuating travel (51) being marginal but somewhat greater. Figure 9 shows an end cap assembly detail, which allows the rotor (3) to rotate in a controlled manner. Also shown is the end seal (16) and its ring spring (26) which forces the end seal (1 ) into contact sealing the end cap face. This figure is a side section of that described earlier in Figures 2 and 3. In this ' figure however, the roller ball bearings can be seen acting between the slide boss (10) and slide bar (8) the exhaust recess by pass (25) on the lower portion of the rotor (3). Figures 10, 10 A and 1 OB show how the pressure motivates the rotor (3), the pressure shown as broad arrows and the exhausting pressure shown as broad dashes. Figure 11 shows a simple scotch crank rotor (3) running through a movement displacement cycle. Figure 11A shows the maximum displacement between the upper volume (12) and the lower (14) these volumes will be referred to retrospectively as the minimum swept volume (12) and maximum swept volume (14). The ports (43 and 44) are set in their prone position, maximum counter rotation set to deliver the maximum delivery. Figure 1 IB shows the rotor (3) inducting through the inlet port (44) and expelling through the outlet port (43). Figure 11C shows the cycle nearing 180° rotational side switchover with the rotor port recess (25) coming into effect allowing the remaining fluid/gas to be expelled through the outlet port (43). Figure 12 shows how the variable delivery port (43,44) plate pump/compressor when double banked can deliver non-fluctuating flow delivery and how the delivery can be substantially reduced by rotational port (43,44) retardation. Figure 12A shows the upper volume of the left hand upper bank rotaiy arrangement expelling into the inducting right hand side of the lower right rotary bank arrangement and vice versa sympathetically annulling any fluctuation between banks. The ports (43.44) are shown in their fully retarded position providing minimum delivery. Figure 12B shows the reduction of volume delivery by overlaying the retarded maximum swept volume (15) over the otherwise maximum swept volume (14) revealing the net delivered swept volume (17) this being less than half on a relatively small rotational retardation of the ports (43 and 44) of around 60°. Figure 13 shows a dual rotor extrusion fitted with steel liners (28). Figure 14 shows a similar arrangement for a dual expansion hot gas, air or vapour rotor (3), with exhaust (44) and inlet ports (43) detailed. The pressure feed link portal (45) is also shown. Figures 15 A, 15B, 15C, 15D and 15E show the dual rotor running through a rotation cycle, where Figure 15F represents primary pressure, Figure 15G represents secondary pressure and finally Figure 15H represents the contracting part of the cycle. Figure 16 shows an insulated heat cell (55) of the preferred format, where the heat block (54) and insulation (58) are encased in steel (59) with an electric heater element (63) fitted in a void, the central cylindrical void is fitted with several heat exchange (40) devices, which are filled and heated by a flame shown later in Figure 19B. Figure 17 shows a heat cell (55) encased in a steel casing (59) with internal insulation (58) an air intake (36) and a heat extraction tube (64) with the heat block (62) hollowed out in the upper portion to provide heat convection flues to enable heat to be extracted by passing cool air from outside through the block which has been heated by the electric heater element (63). In Figure 18 A, the lower most element (60) shows an arrangement, by which the element (60) is also hollow, however, in this case the tube (64) filled with heat fluid (19) such as oil. The heat in this case is conveyed by flow convection or otherwise. The bulbous portion and subsequent piping are also insulated (58) up to the point of use. Figure 18B shows varied forms of heat extraction elements (60) to enable heat to be extracted, conveyed and utilised. The element (60) is a solid profile that would conduct heat through its large flat surface area through the bulbous rod-like portion, made of a conductive material such as copper. Figure 19 shows a hot gas rotor (53) the main component consisting of a twin bank of double expansion rotors. Mounted either side of the main bank are two heater chambers (39) internally fitted with heat exchangers (40). The air is draw in through air filters (37) by air compressors (38) and passed over the heat exchanger (40). The pressurised hot air is then passed into and drives the double expansion rotor (3) as detailed in Figures 15, 15A, 15B, 15C, 15D and 15E. The expanded and cool gases exhaust out of the pipes (54) slung either side of the main bank. Figure 19 A shows the heat exchanger (40) that is a flattened tube wound into a clock spring configuration with a small air space between coils. Figure 19B shows the hot gas rotor arrangement (53) and the power source in the form of a flame (52) charged heat cell (55) that could additionally be charged by an electrical element. The heat cell (55) would be in the form of that shown in Figure 16, the flame (52) could be generated from combustible gas or liquid fuels depending on the nozzle (41) type. The heat exchangers (40) are heated by the flame (52) the heat exchanger (40) tubing is oil (19) filled to pipe the heat to the on board heat exchangers (40) within the heater chambers (39) of the hot gas rotor. Figure 20 shows a graph that demonstrates the large difference between the efficiency levels of heat engines and internal combustion engines with the more efficient heat engine at 50% on the left and the internal combustion engine on the right. Figure 21 shows the dramatic difference of power to weight ratios in comparing a magnetite heat cell (55) with a conventional rechargeable vehicle battery with the magnetite heat cell's (55) higher energy content on the left and the lower rechargeable vehicle battery on the right. Figure 22 shows a pump known as a gyrotor. The rotor (3) is four-lobed and is mounted on a shaft, meshing and rotating in the same direction as the outer annular ring profile (70). The arrangement creates a device with four active chambers, displacing as it rotates. The ports (43,44) are shown by a dashed line and can be rotated or rotationally distorted in the same manner as shown in figure 12 to vary the displacement and delivery of the pump.

Claims: CLAIMS 1. A rotary displacement geometry for use as a motor, pump or compressor, in which the simple motion of a two sided rotor (3) sliding through a displaced axial centre creates compression ratios of between two and fifteen to one, at maximum displacement, the chamber profile (1) being a loop which could be circular, part circular or folly non-circular, in each case the profile (1) will be contacted by either the two longitudinal rotor edges or seals (5) mounted there upon, the axial displaced centre being offset from the actual mid-centre (33) by less than one-sixth of the chambers (1) effective internal diameter (2), and the rotor's (3) circumscribed foil cross sectional area being 40% or more of the chamber's (1) foil cross sectional area. 2. A rotary displacement geometry according to claim 1 , in which the rotor (3 ) element traverses the majority of sliding motion with seals (5) having a directional slide travel of one-twentieth or less of the chamber's effective internal diameter. 3. A rotary displacement geometry according to any one of the preceding claims, in which the rotor's sliding travel is governed and restrained by bearings (7) set in a peripheral track (11) arrangement at either end of the rotor assembly, and the bearings (7) run upon a peripheral track (11) corresponding to the chamber (1) or a set radial distance away. 4. A rotary displacement geometry according to any one of the preceding claims, in which the chamber ( 1 ) periphery follows a dedicated geometry whereby the reduced volume chamber (1) portion below the displaced centre has a non-circular profile taking the form of a parabola or half ellipse split along the long axis, the extended volume chamber portion in both cases being derived from the reduced volume chamber portion, with a locus point tracked at a rotor width distance (2) traversing the displaced centre. 5. A rotary displacement geometry according to any one of the preceding claims, in which the chamber (1) profile is in part or fully non-symmetrical about any angle through the displacement centre, whereupon the profile is sufficiently distorted to dispose the rotor's (3) radial lever arm fully for up to 130° of rotational travel. 6. A rotary displacement geometry according to any one of the preceding claims, in which the chamber profile incorporates a portion of part circular geometry that by 5 definition being that, which follows a set distance from one point or centre. 7. A rotary displacement geometry according to any one of the preceding claims, in which the sliding motion of the rotor (3) is directly in sliding contact with the chamber (1) in the more basic form. 8. A rotary displacement geometry according to any one of claims 1, 4, 5, 6 and 7, 10 where the main shaft (68) provides the sliding track, by flat face, or roller contact with the rotor (3) running as a scotch crank sliding on the shaft (68) surface. 9. A rotary displacement geometry according to any one of the preceding claims, in which either inlet (44) or exhaust (43) ports are fed into the end or mid cap plates of the rotary multi-lobed displacement units, and where upon the ports (43 , 44) may 15 be rotated or distorted in radial arcs to retard displacement and reduce delivery. 10. A rotary displacement geometry according to any one of the preceding claims, in which the edges or ends of the rotor (3) are fitted with seals (16) 11. A rotary displacement geometry according to any one of the preceding claims, in which the rotary device is configured as a vapour or hot gas engine. 20 12. A rotary displacement geometry according to any one of the preceding claims, in which the devices are configured as a rotary internal combustion engine. 13. A rotary displacement geometry according to any one of the preceding claims, in which the rotary devices are configured as a pressure charged internal combustion engine. 25 14. A rotary displacement geometry according to any one of claims 1 and 10, where the rotary device takes the form of a heat engine powered by an external burner and or heat cell (55). 15. A rotary displacement geometry according to any one of the preceding claims 1, 10 and 13, where the heat engines source is either directly by flame (52) or by piped hot fluid (19). 16. A rotary displacement geometry according to any one of the preceding claims 1, 10, 5 13 and 14, where the heat source is a thermal storage ceil (55) with an insulated jacket (58). 17. A rotary displacement geometry according to any one of the preceding claims, in which the rotor (3) and or chamber (1) is made of metal, in the form of extrusion casting or deformed tube. 10 18. A rotary displacement geometry according to any one of the preceding claims, in which the rotor (3) or chamber (1) is non-metallic, in the form of ceramics or plastic based material. 19. A rotary displacement geometry according to any one of the preceding claims, in which the rotor (3) arrangement is configured as a heat engine (53) where the 15 inducted gas or air, inert or otherwise is drawn in and pre-compressed by fan or compressor (38) prior to heating or combustion. 20. A rotary drive mechanism for use as a motor, a pump or a compressor, the mechanism including a housing having a chamber therein defined by a peripheral wall, a rotor rotatably mounted in the chamber having two longitudinal seal edges 20 that contact the profiled wall, the rotor being mounted for rotation about a rotation axis and for sliding movement relative to the rotation axis in a direction perpendicular thereto, the rotation axis being offset from the centre of the chamber. 21. A rotary drive mechanism according to claim 20 including a rotary displacement geometry according to any one of claims 1 to 19. 25 22. A motor, a pump or compressor including a rotary drive mechanism according to claim 20 or claim 21. 23. A variable delivery pump mechanism including a housing having a chamber therein defined by a peripheral wall, a rotatable rotor mounted in the chamber and dividing said chamber into a plurality of sub-chambers, the rotor being constructed and arranged such that rotation thereof causes the volume of said sub-chambers to increase and decrease alternately, and an inlet/outlet port member having at least one inlet port and at least one outlet port therein, said inlet/outlet port member being rotatably adjustable relative to the housing, to adjust the delivery volume of the pump.

Inventors: OLLIS WILLIAM HENRY, OLLIS, WILLIAM, HENRY
Applicants: OLLIS WILLIAM HENRY, OLLIS, WILLIAM, HENRY
Application year: 2001
Main classification: F01C1
IPC codes: F01C1