Patent Grant
6877476
The present invention relates to internal combustion engines. More particularly, the present invention relates to an internal combustion engine which utilizes a pinwheel engaged with a slotted disc to define the compression chambers of the internal combustion engine.
The transportation industry, as well as other numerous industries, rely on the internal combustion engine as a source of power. The typical internal combustion engine employs one or more cylinders having pistons reciprocating therein. The reciprocating motion of the pistons defines an intake stroke, a compression stroke, a power stroke, and an exhaust stroke. A mixture of air and fuel is directed into the cylinder through an intake valve during the intake stroke; the air and fuel mixture is compressed during the compression stroke; the air and fuel mixture is ignited and burns during the power stroke; and, finally, the exhaust gases are exhausted from the cylinder through an exhaust valve during the exhaust stroke.
Engine speed, power, fuel consumption and emissions are typically controlled by manipulating various parameters associated with the engine. These parameters include, but are not limited to, fuel injection pressure, ignition timing, induction throttling, exhaust gas recirculation, air injection, displacement on demand systems, and induction charging. Various sensors monitor the operating conditions of the engine and these sensors provide feedback to an engine control unit (ECU) that, in turn, adjusts the parameters associated with the engine.
Typical engine configurations include V-shaped engines, in-line engines, 2-cycle engines, 4-cycle engines, gasoline engines, natural gas engines, diesel engines, and the like. While each of these engines or combination of engines has proven performance in today's applications, the continued development of internal combustion engines includes the development of engines that have improved performance, reliability and efficiency with lower emissions, and preferably have fewer components and are less costly to produce.
The present invention provides the art with an engine configuration that can operate over a broad-speed range; it can take advantage of the Atkinson cycle; it has all rotary moving parts; and it has a combustion chamber having a low surface-to-volume ratio. The present invention also provides a compact, lightweight engine that can accommodate gas direct injection (GDI) and allows for adjustable injector positioning.
The engine of the present invention utilizes a pinwheel engaged within a slot formed in a disc. The pinwheel and the disc rotate together such that the pin of the pinwheel moves laterally within the slot to define the compression chamber. The engine delivers improved performance, reliability and efficiency with lower emissions than the current state-of-the-art engine configurations.
Further areas of applicability of the present invention will become apparent from the detailed description provided hereinafter. It should be understood that the detailed description and specific examples, while indicating the preferred embodiment of the invention, are intended for purposes of illustration only and are not intended to limit the scope of the invention.
The present invention will become more fully understood from the detailed description and the accompanying drawings, wherein:
The following description of the preferred embodiment(s) is merely exemplary in nature and is in no way intended to limit the invention, its application, or uses.
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The rotation of indexer 16 and pinwheels 14 causes opposing slots 72 of indexer 16 to receive an initial charge of air from intake ports 40 and 48. This equates to an intake stroke. The initial charge of air can be at atmospheric pressure or it can be pressurized by a turbo-charger, a blower or by other means known in the art. Continued rotation of indexer 16 and pinwheels 14 will cause a piston pin 62 from each pinwheel 14 to enter one of the opposing slots 72. As each piston pin 62 enters its respective slot 72 and each pin 62 sealingly engages the wall of its respective slot 72, the outer surface of indexer 16 passes the sealing portion of each sealing member 82 and each opposing slot 72 becomes a sealed compression chamber 92.
Continued rotation of indexer 16 and pinwheels 14 will cause each piston pin 62 to traverse further and further into its respective slot 72 until the position shown in
Engine 10, as illustrated, has four piston pins 62 on each pinwheel 14 and four slots 72 in indexer 16. Thus, a single revolution of each pinwheel 14 results in all four piston pins 62 of each pinwheel 14 completing all four strokes of the power cycle. Thus, engine 10 delivers four power strokes per pinwheel 14 per revolution as compared to one power stroke per revolution for a prior art 4-cylinder, 4-stroke reciprocating engine. Therefore, engine 10 having a displacement of 106 cc can deliver the same power at the same speed as an 848 cc 4-cylinder, 4-stroke reciprocating engine. Alternately, engine 10 can deliver the same power as the prior art 4-cylinder, 4-stroke reciprocating engine, but at ⅛ of the engine speed. Also, because of the multiple power pulses per revolution and the inherent balance of engine 10, a broad speed range between an ultra low idle (125 rpm) and a high top speed (8000 rpm) should be attainable.
The following presents various advantages of the engine of the present invention over the prior art 4-cylinder, 4-stroke reciprocating engine.
To maximize the amount of energy converted from fuel to thrust, the least amount of heat should be allowed to escape through the combustion chamber walls. If heat escape can be minimized, then more heat will stay in the gas and the peak pressure will be greater. The ratio of surface area to the volume in the chamber at TDC is a measure of thermodynamic efficiency. The larger the surface area for any given volume, the larger the heat losses. Table 1 shows a comparative analysis of the engine of the present invention compared with an identically sized reciprocating engine with the same compression ratio (cr). In this comparison, the reciprocating model was calculated assuming a pancake-style cylinder head with no squish/quench zones. The engine of the present invention has a large squish zone and its surface area is included in the calculation.
A better s/v ratio will reduce HydroCarbon (HC) emissions. HC is released to the exhaust when the flame cools and stops burning as it reaches the chamber walls. The lesser the surface area, the less HC will be released.
The engine of the present invention delivers better s/v due to the fact that there are, effectively, two chambers joined together back-to-back. Because there is no back to the tunnel, its surface area is reduced and its volume increased. Also, the overall trumpet shape of the chamber is closer to an ideal s/v ratio (a sphere) than the pancake chamber used in many reciprocating engines. Thus, an 8.3% improvement in s/v ratio is likely to translate into increased peak pressure and decreased HC release.
Pre-ignition. Pre-ignition occurs when a hot spot ignites the mixture ahead of the timed spark. This advances combustion so that too great a pressure build-up occurs before TDC. This creates excessive negative work. Pre-ignition can be caused by some overheated protruding part in the chamber. The spark plugs, sharp corners, direct injectors or carbon deposits can be a cause of hot spots in both the engine of the present invention and reciprocating engines, but only reciprocating engines can get hot spots from valves and gaskets, as the engine of the present invention has neither.
Detonation. When the end mixture advances in front of the flame front in a non-turbulent manner, it gets backed into a corner where the pressure builds up to such an extent that it explodes spontaneously. This is detonation. Abnormally high pressure waves ring through the chamber giving a characteristic pinging or knocking sound and can damage the engine. Detonation scours the protective boundary layer and oil film causing further deterioration. A slow-moving, non-turbulent flame front and lean mixtures is the leading cause of detonation. The engine of the present invention has four power strokes per revolution and, therefore, has the potential to idle below 125 rpm. At such a low speed, detonation is more likely. Boosted compression ratio also increases the likelihood of detonation but, fortunately, GDI, stratified charge and excess EGR decreases detonation tendency. Thus, pre-ignition is less likely in the engine of the present invention, as the chamber has fewer sharp edges than the reciprocating engine where hot-spotting can occur. Detonation would be a serious problem at ultra low speeds if it weren't for the use of GDI and stratified charge technology.
Duel ignition reduces the flame path and allows for less advanced ignition timing. In comparing reciprocating engines with either single or twin spark plugs, duel ignition can complete the burn time 6′ (of crank angle) sooner. This generates (i) a higher peak pressure (more torque), (ii) less negative (compressive) work with less advance, (iii) less cyclic dispersion, (iv) improved fuel economy and (v) improved response; all of these, particularly at mid to low loads. The engine of the present invention uses duel ignition and should mirror the twin plug reciprocating engine's advantages. Additionally, the engine of the present invention has GDI but with two orifices on each injector tip, each of which is aimed at a spark plug. Thus, two separate stratified charges are generated at low load.
Twin plugs are necessary in the engine of the present invention to generate performance advantages at low and mid speed ranges and to facilitate full advantage of the back-to-back, tunneled combustion chamber.
In Table 2 below, we see the data for the same identical engine sizes and compression ratio. The reciprocating engine begins compression at bottom dead center (BDC). Here, we assume the pressure is at atmospheric, 1 bar. If this engine is motored (no combustion), then peak pressure at TDC is 22.4 bar (adiabatic compression). After TDC (aTDC), the pressure begins dropping (13.6 bar) as the piston begins its expansion stroke. At BDC, expansion ends with pressure returning to atmospheric (reversible adiabatic), 1 bar. When a theoretical 60 bar is added (pressure generated from fuel combustion), then pressure peaks at 73.6 bar aTDC. This pressure thrusts the piston to BDC. However, at BDC, there is still 5.4 bar of pressure left over that will be wasted when the exhaust valve opens.
The engine of the present invention benefits from starting its compression stroke before BDC (bBDC). Thus, at BDC, the pressure has risen to 1.3 bar, getting a 0.3 bar head start (+30%), or boost, on the reciprocating engine. By the time the piston pin has reached TDC (in the motored case), the pressure peaks at 28.6 bar compared to 22.4 bar for the reciprocating engine (adiabatic compression). At BDC, the pressure returns to 1.3 bar. But, because there is addition expansion capability, it can further expand to 0.7 bar (a partial vacuum). When 60 bar is then added (pressure generated from fuel combustion), pressure peaks at 77.3 bar aTDC. This pressure thrusts the piston pin through BDC (where it is at 5.7 bar) and continues thrusting the pin onward until the exhaust port is exposed (aBDC). Here the pressure is reduced to 3.0 bar, which is then wasted down the exhaust system. This extended expansion is known as the “Atkinson cycle.”
In
Clearly, there is some additional negative work for the engine of the present invention in compressing the extra inducted air but there is significantly more positive work output due to the Atkinson expansion. Table 3 shows the Work (J) comparison.
Thus, as shown in Table 3, the engine of the present invention outperforms the reciprocating engine by 14% when comparing an equivalent single slot on the engine of the present invention versus a single cylinder on the reciprocating engine. This net 14% gain is the residual work due to the Atkinson expansion less the negative work for the additional (boosted) compression. Thus, all else being equal, the engine of the present invention is 14% more powerful or delivers 14% less specific fuel consumption (sfc) than the reciprocating engine. Of particular note is the capability of the engine of the present invention to generate power. Unlike the 4-cylinder reciprocating engine which has one power stroke per revolution, the engine of the present invention has eight. Thus, the power per revolution is 9.12 times (8×1.14) as great as the equivalent reciprocating engine. With this range of power, it may be possible to negate the necessity for gearboxes in some applications. Note, it would still require eight times as much fuel to generate eight times the power, but the net 14% “Atkinson dividend” is still a direct sfc or power advantage. These comparisons indicate the theoretical work and power output; they do not take into consideration other losses, such as friction or pressure leakage.
We have identified the additional compression boost that is available with the engine of the present invention without the need for a super-charger or turbo-charger. However, the analysis made thus far assumed an identical energy input (60 bar) from fuel. With the additional 0.3 bar of boost (+30%) over atmospheric achieved in Table 2, it is possible to add an additional 30% fuel and still achieve a stoichiometric mixture. Table 4 shows the impact of this additional energy input.
Table 4 is similar to Table 2,
Thus, the engine of the present invention has potential to generate 47% more work when boosted with commensurate fuel input than an equivalent reciprocating engine. Power per revolution is 11.7 times (8×1.47) as great as the equivalent reciprocating engine. Note: these comparisons indicate the theoretical work and power output; they do not take into consideration other losses, such as friction or pressure leakage.
From Blackmore and Thomas (1977), an analysis of a 1.5 liter reciprocating engine operating at 4000 rpm generated losses as follows:
Using this analysis as a benchmark, we can estimate the friction and pumping losses associated with the engine of the present invention (assumptions in parenthesis):
Thus, the engine of the present invention delivers a net power gain of 6.6 kW at 4000 rpm. Friction and pumping losses are reduced by an estimated 58%.
In most SI engines, a throttle is used to control the amount of air entering the cylinder. Fuel is injected either near the throttle body (single-point fuel injection —SPFI) or near each inlet port (multi-point fuel injection—MPFI). The timed pulse of the injection event is controlled by the ECU to deliver a (close to) stoichiometric mixture to the cylinder. At low load, less fuel and less air (more throttle) enter the cylinder. At high load, the throttle is opened wide (WOT) and a longer pulse of fuel is injected into the port. In either case, a homogeneous mixture enters the cylinder.
In a gas direct injection (GDI) engine, the injector sprays fuel directly into the cylinder. At low load, it is desirable to spray the fuel into the vicinity of the spark plug, creating a moving cloud of (ideally) stoichiometric mixture that will be ignited by the spark plug at the right instant. Outside of this cloud, the gas (air and/or exhaust gas residue—EGR) has little or no fuel in it. Therefore, only a small combustion event occurs, providing reduced thrust on the piston. Low load combustion can thus be achieved without the need to throttle the incoming air. This significantly reduces pumping losses. Additionally, during this stratified mode, it is impossible for detonation to occur, as there is no surrounding fuel/air mixture that can spontaneously ignite. Thus, a higher compression ratio can be tolerated. Typically, GDI engines deliver 10-15% specific fuel consumption (sfc) performance improvements of PFI engines.
At higher loads, it is desirable to spray the fuel centrally into the cylinder in order to create a more homogeneous mixture throughout. However, until now, a compromise has been needed to position the injector to achieve both high load homogeneous cylinder filling and stratified cloud formation at the spark plug. This compromise has normally been met by utilizing the cylinder wall or piston crown as a mechanism to deflect the spray cloud to the spark plug (when near TDC). The spray impingement on the wall or piston, unfortunately, causes poor emissions and/or burning performance and/or causes the formation of deposits.
In the engine of the present invention, the injector has two orifices, each pointing to a diametrically opposed spark plug at each end of the combustion tunnel. The injector is held stationary while the indexer moves about it on the same axis. However, the injector can be rotated back and forth about some angle such that, in low rpm, stratified mode, the injection event will occur such that the two spray clouds will reach the spark plugs at the optimum time. In higher speed, homogeneous mode, the injector can be rotated (under ECU command) to ensure that the spray is directed more down the center of the cylinder. In fact, at higher speeds and longer injection durations, the spray pattern becomes fanned out, more completely creating homogeneity. This directable injection system is termed active direction injection (ADI).
The relatively long distance for the sprays to reach the spark plugs in the combustion tunnel of the engine of the present invention provides better mixture preparation and combustion than short-throw or wall-guided configurations.
Thus, GDI delivers 10-15% reduced sfc. In the engine of the present invention, the proximity to and directability of the ADI system is predicted to provide better stratified, transitional and homogeneous mode performance with reduced HC emissions.
In reciprocating engines, there exist primary and secondary motions of the piston which cause vibration and fatigue, as well as couple moments that create torsional vibration in the crankshaft. These forces can be balanced by crank geometry, cylinder arrangement (e.g, in-line 6, V8, etc.) and by using countermeasures such as dampers and counter-rotating masses. However, all these countermeasures add weight and complexity.
In the engine of the present invention, the pinwheels are in perfect balance and only the indexer undergoes sinusoidal motion about its axis. There are no couple moments. By using two pinwheels and tunnel chambers on the indexer, the combustion forces on the indexer shaft are also completely balanced.
The sinusoidal forces causing the angular velocity on the indexer to speed up and slow down will need to be absorbed within the indexer itself. Thus, reducing its weight (and, thus, inertia) is critical in minimizing these forces.
In a reciprocating engine, the continuous pulsing (once per revolution in a 4-stroke, 4-cylinder) caused by the combustion needs to be smoothed out by a heavy flywheel. In the engine of the present invention, the pulsing occurs every ¼ of a revolution. So at a comparable speed, the engine of the present invention requires less pulse smoothing. In fact, the pinwheels themselves act as flywheels to help provide smooth operation.
A current trend to reduce sfc at low idle and cruise conditions is to deactivate the fuel and/or ignition and/or valve actuation of specified cylinders in 6- to 12-cylinder engines. The deactivation is generally rotated about the cylinders so as to maintain thermal and tribological equilibrium. The deactivation also allows a wider throttle opening and, thus, higher efficiency operation in the remaining active cylinders. DOD is not typically conducted in 4- or 6-cylinder engines, as balance is compromised.
In the twin pinwheel of the engine of the present invention, DOD is possible by deactivating the fuel and spark from a tunnel chamber. Although GDI technology in itself has the potential to eliminate the necessity for DOD, the ultra low rpm capability may make DOD attractive for some applications. The two tunnels, labeled, say, X and O may be deactivated in rotation as follows:
O X O X O X X O X O X O
O X O X X O X O O X O X
O X X O O X X O O X
Clearly balance is maintained for smooth operation. However, (negative) work is required to compress the inducted air (partially offset by positive work as it expands past TDC), as the engine of the present invention has no valves it can hold open. Thus, DOD can deliver up to 33% sfc savings.
In order to reduce CO and HC emissions after combustion, additional air can be introduced into the exhaust stream to help extend combustion. About 20% injected air gives a good balance of CO and HC reduction in homogeneous mode operation. In reciprocating engines, an auxiliary air pump system is required.
Inherently, the engine of the present invention has its own integrated, automatic air pumping system, albeit at a fixed volume. When the piston pin enters the slot, the space immediate behind the piston pin contains a slug of fresh air. As the piston pin travels down the slot, this slug of air follows it. After TDC, when the piston pin begins its expansion stroke, it now thrusts this slug of fresh air out into the exhaust stream of the previous pin's exhaust stroke. Thus, each slug of exhaust is chased by a slug of fresh air. Any remaining HCs have added oxygen to continue their burning.
During cold weather starts, an enriched mixture is required. With the integrated air injection of the engine of the present invention, more burning is likely to take place down the exhaust tract, which will assist in getting the catalyst up to operating temperature (fast light-off). Thus, air injection is achieved at no added cost or complexity, resulting in lower CO and HC emissions.
Clearly, there are fewer fixed and moving parts in the engine of the present invention. Consider also that this compares an 8-slot engine of the present invention with a 4-cylinder reciprocating engine. Examining the principal components in Table 6 below:
All else being equal, the engine of the present invention has only 46% of the principal parts of a reciprocating 4-cylinder engine. This translates into higher reliability and lower costs.
The over-riding impact on emissions is the ability of the engine of the present invention to significantly reduce sfc. A large engine certified as ULEV still produces more emissions than a smaller ULEV engine.
The extended Atkinson cycle, stratified charge, and DOD at low loads and induction boosting at higher loads are predicted to provide significant sfc savings. This directly translates into lower carbon dioxide (CO2) emissions, the main greenhouse gas.
A smooth combustion chamber and piston pin design (with few crevices) improves thermodynamic efficiency and is predicted to reduce HC and Co.
NOx is reduced by EGR in homogeneous mode, but has the potential to be higher in stratified mode unless post exhaust system strategies are employed.
Residual HC is burnt with air injection, thereby further reducing CO and CO2.
The combination of advantages of the engine of the present invention is predicted to significantly reduce all emissions regardless of fuel type.
The engine of the present invention rivals a boxer engine in terms of compactness. At an overall height of under 150 mm (length 975 mm; width 490 mm), it is flatter per liter of capacity than any IC engine available. Volume envelope, before ancillaries, is typically 60×106 mm3, which is comparable to an in-line 4-cylinder sized 660 mm×460 mm×200 mm.
Compactness provides a very low center of gravity, aerodynamic and packaging advantages. All of which translate into better vehicle dynamics, performance and safety.
In the engine of the present invention, the basic 1696 cc weight is 87.5 kg. This compares favorably to VW's air cooled flat four 1600 cc at 91 kg, Honda's liquid cooled 1488 cc at 102 kg and Subaru's liquid cooled boxer 1781 cc at 97 kg (all in aluminum).
Referring now to
Outer housing 112 defines a first cylindrical wall 130, a second cylindrical wall 132 and a third cylindrical wall 134. First cylindrical wall 130 defines a first cavity 136 within which is positioned one of pinwheels 114. A bearing 138 supports pinwheel 114 for rotation with the first cavity 136. A first intake port 140 and a first exhaust port 142 extend through the outer housing 112. Second cylindrical wall 132 defines a second cavity 144 within which is positioned one of pinwheels 114. A bearing 146 supports pinwheel 114 for rotation with the second cavity 144. A second intake port 148 and a second exhaust port 150 extend through the outer housing 112. Third cylindrical wall 134 defines a third cavity 152 within which is positioned indexer 116. A bearing 154 supports indexer 116 for rotation with the third cavity 152.
The pair of pinwheels 114 each comprise a circular disc 160 and a plurality of piston pins 162 (four illustrated in
Indexer 116 comprises a disc-shaped member 170 defining a plurality of slots 172 (four illustrated in
Pinwheels 114 are assembled to indexer 116 with each pinwheel 114 having one of the plurality of piston pins 162 disposed within a respective slot 172 (only one set of piston pins shown in FIGS. 15 and 16). Piston pins 162 of pinwheels 114 are disposed in opposing slots 172 at a position which would equate with top dead center (TDC) of a typical piston internal combustion engine. A continuous synchronizing belt 182 has a plurality of teeth which mate with a plurality of teeth formed on each pinwheel 114 to synchronize the rotation of the two pinwheels 114. A pair of belt tensioners 184 located on opposite sides of pinwheels 114 maintain tension on belt 182. An axle shaft 186 extends through outer housing 112 to rotatably support each pinwheel 114. A seal 188 is disposed around each slot 172 on both sides of indexer 116. When indexer 116 is assembled with the pair of pinwheels 114, seals 188 sealingly engage circular disc 160 of each pinwheel 114 and sealingly engage the bottom surface of outer housing 112 to isolate each slot 172.
One pinwheel 114 is disposed within first cavity 136, the second pinwheel 114 is disposed within second cavity 144, and indexer 116 is disposed within third cavity 152. Each pinwheel 114 has one piston pin 162 disposed within a respective slot 172 of indexer 116 as shown in
The rotation of indexer 116 and pinwheels 114 causes opposing slots 172 of indexer 116 to receive an initial charge of air from intake ports 140 and 148. This equates to an intake stroke. The initial charge of air can be at atmospheric pressure or it can be pressurized by a turbo-charger, a blower or by other means known in the art. Continued rotation of indexer 116 and pinwheels 114 will cause a piston pin 162 from each pinwheel 114 to enter one of the opposing slots 172.
Continued rotation of indexer 116 and pinwheels 114 will cause each piston pin 162 to traverse further and further into its respective slot 172 until it reaches its inner most position within its respective slot 172. This causes compression of the initial air charge within the sealed compression chamber and equates to a compression stroke. At a pre-selected time during this compression stroke of piston pin 162, one fuel injector 176 of fuel injection system 118 injects fuel into its respective tunnel 178, which injects fuel into each of the opposing slots 172 or each of the opposing compression chambers. Fuel injectors 176 include two jet orifices 194 with one jet orifice 194 being directed toward a respective tunnel 178. At low load, one orifice is pointed toward spark plug 180 in one slot 172 and the other orifice 194 is pointed to spark plug 180 in the opposing slot 172 for stratified charge combustion. At higher loads, injectors 176 can be rotated by a small electric motor (not shown) such that a fan-like spray pattern can be mapped for (medium to full load) whole compression chamber filling and combustion. At a pre-selected time just prior to or at a pre-selected time just after each piston pin 162 reaches its inner most position, which is equivalent to top dead center in the prior art engines, the two spark plugs 180 of ignition system 120, which are located within the opposing slots 172 or the opposing compression chambers are activated to ignite the fuel-air mixture that is within each opposing slot 172 or the compression chamber and its associated tunnel 178. The ignition and subsequent burning of the fuel-air mixture within each slot 172 or the compression chamber causes continued rotation of indexer 116 and pinwheels 114. This is equivalent to a power stroke. The continued rotation of indexer 116 and pinwheels 114 will cause each piston pin 162 to traverse further and further out of its respective slot 172 and piston pins 162 will exit their respective slots 172; and these slots 172 will align with exhaust ports 142 and 150 and into exhaust passages 124 to exhaust the products of combustion created by the ignition and burning of the fuel-air mixture. This is equivalent to an exhaust stroke. Continued rotation of indexer 116 and pinwheels 114 will cause the opposing slots 172 to receive an additional air charge and the cycle begins again.
Engine 110, as illustrated, has four piston pins 162 on each pinwheel 114 and four slots 172 in indexer 116. Thus, a single revolution of each pinwheel 114 results in all four piston pins 162 of each pinwheel 114 completing all four strokes of the power cycle. Thus, engine 110, similar to engine 10, delivers four power strokes per pinwheel 114 per revolution as compared to one power stroke per revolution for a prior art 4-cylinder, 4-stroke reciprocating engine. This provides the same features and advantages as described above for engine 10.
Referring now to
The present invention provides the art with an invention that greatly improves performance with low fuel consumption and emissions. Some, but not all, of the advantages of the engine of the present invention are:
The description of the invention is merely exemplary in nature and, thus, variations that do not depart from the gist of the invention are intended to be within the scope of the invention. Such variations are not to be regarded as a departure from the spirit and scope of the invention.
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