EP0057591A2 - Brennkraftmaschine - Google Patents

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Publication number
EP0057591A2
EP0057591A2 EP82300475A EP82300475A EP0057591A2 EP 0057591 A2 EP0057591 A2 EP 0057591A2 EP 82300475 A EP82300475 A EP 82300475A EP 82300475 A EP82300475 A EP 82300475A EP 0057591 A2 EP0057591 A2 EP 0057591A2
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EP
European Patent Office
Prior art keywords
chamber
cylinder
power
engine
compressor
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Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Granted
Application number
EP82300475A
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English (en)
French (fr)
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EP0057591B1 (de
EP0057591A3 (en
Inventor
Clyde C. Bryant
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Individual
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Individual
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Publication of EP0057591A3 publication Critical patent/EP0057591A3/en
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B41/00—Engines characterised by special means for improving conversion of heat or pressure energy into mechanical power
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/02—Engines with reciprocating-piston pumps; Engines with crankcase pumps
    • F02B33/06—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps
    • F02B33/22—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps with pumping cylinder situated at side of working cylinder, e.g. the cylinders being parallel
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/02—Engines with reciprocating-piston pumps; Engines with crankcase pumps
    • F02B33/28—Component parts, details or accessories of crankcase pumps, not provided for in, or of interest apart from, subgroups F02B33/02 - F02B33/26
    • F02B33/30—Control of inlet or outlet ports
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/32—Engines with pumps other than of reciprocating-piston type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/44—Passages conducting the charge from the pump to the engine inlet, e.g. reservoirs
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/16—Engines characterised by number of cylinders, e.g. single-cylinder engines
    • F02B75/18—Multi-cylinder engines
    • F02B75/20—Multi-cylinder engines with cylinders all in one line
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B1/00—Engines characterised by fuel-air mixture compression
    • F02B1/02—Engines characterised by fuel-air mixture compression with positive ignition
    • F02B1/04—Engines characterised by fuel-air mixture compression with positive ignition with fuel-air mixture admission into cylinder
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/16—Engines characterised by number of cylinders, e.g. single-cylinder engines
    • F02B75/18—Multi-cylinder engines
    • F02B2075/1804—Number of cylinders
    • F02B2075/1812—Number of cylinders three
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/16—Engines characterised by number of cylinders, e.g. single-cylinder engines
    • F02B75/18—Multi-cylinder engines
    • F02B2075/1804—Number of cylinders
    • F02B2075/182—Number of cylinders five
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B2275/00—Other engines, components or details, not provided for in other groups of this subclass
    • F02B2275/34—Lateral camshaft position
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B3/00—Engines characterised by air compression and subsequent fuel addition
    • F02B3/06—Engines characterised by air compression and subsequent fuel addition with compression ignition

Definitions

  • This invention relates to a method of deriving mechanical work from combustion gas in an internal combustion engine by means of a new thermodynamic working cycle and to reciprocating internal combustion engines for carrying out the method.
  • thermodynamic efficiency increases as the expansion ratio of an internal combustion engine is increased, more energy is extracted from the combustion gases and the thermodynamic efficiency increases. It is further understood that increasing compression increases both power and fuel economy due to further thermodynamic improvements.
  • the objectives for an efficient engine are to provide high compression, begin combustion at maximum compression and then expand the gases as far as possible against a piston.
  • Rotary engines have also been patented which strive to gain the same advantages.
  • One such is the new Wankel engine, U.S Patent No. 3,688,749 issued in 1972, in which a charge is compressed in one chamber of the rotor of a four-lobed rotor engine where the charge is ignited and expanded first in the initial chamber and then through a duct into the next down-stream chamber.
  • Some of the problems with this concept are that the second expansion chamber is already half filled with recompressed exhausted gases from the previous firing and there are extensive throttling losses in transferring the charges.
  • the present invention provides a reciprocating internal combustion engine comprising a compressor chamber for compressing an air charge, power chambers in which combustion gas is ignited and expanded, a piston operable in each chamber and connected to a crankshaft by connecting link means for rotating the crankshaft in response to reciprocation of each piston, a transfer manifold communicating said compressor chamber with said power chambers through which manifold the compressed charge is transferred to enter the power chambers, an admission valve controlling admission of air to said compressor chamber for compression therein, an outlet valve controlling admission of the compressed charge from the compressor chamber to the transfer manifold, an intake valve controlling admission of the compressed charge from the transfer manifold to said power chambers, and an exhaust valve controlling discharge of the exhaust gases from said power chambers, said valves being timed to operate such that the air charge is maintained within the transfer manifold and introduced into the power chamber without any appreciable drop in charge pressure so that ignition can commence at substantially maximum compression, means being provided for causing fuel to be mixed with the air charge to produce a combustible gas, means being provided for ignition of
  • the compression ratio for spark ignited engines can be increased without the attendant problem of combustion detonation, the expansion ratio for both spark ignited and compression ignited engines is greatly increased, and a much greater charge turbulence is produced in the combustion chamber of both.
  • crankshaft The extra power strokes per revolution of crankshaft translates into a nominal 2-2/3 stroke cycle engine in the tor 8-cylinder design and produces a nominal 3-stroke cycle engine in the 3- or 6- cylinder design for reduced friction and greater mechanical efficiency.
  • Figure 1 shows a four cylinder reciprocating internal combustion engine for gasoline, diesel, gas or hybrid dual-fuel operation and having four cylinders 2-5 in which pistons 6-9 respectively are arranged to reciprocate.
  • Pistons 6 - 9 are connected to a common crankshaft 10 in conventional manner by means of connecting rods 11-14, respectively.
  • Engine 1 is adapted to operate in a 2-stroke cycle so as to produce three power strokes per revolution of the crankshaft 10.
  • one cylinder 5 functions as a compressor, so that during operation of the engine, compressor cylinder 5 takes in an air charge at atmospheric pressure, or alternatively an air charge which previously has been subjected to suoercharging to a higher pressure, via an admission control valve 'a', through an intake conduit 15.
  • Manifold 16 is constructed and arranged to distribute the com- pressued charge by means of branch conduits 17, 18 and 19 and intake valves 'i' to the three remaining (expander) cylinders 2, 3 and 4 respectively which produce the power of the engine.
  • the volume of the combustion chamber of each expander cylinder 2, 3 and 4 is preferably sized to be no larger than one third that of a conventional engine having a similar compression ratio. This is because the total volume of the combustion chambers should not exceed the volume of charge compressed by the compressor piston and therefore no expansion of the gases will occur before combustion takes place.
  • Engine 1 has a camshaft 20 which is arranged to be driven at the same speed as the crankshaft in order to supply one working stroke per revolution for both power and compressor pistons, as described hereinafter.
  • the piston At the time the intake valve opens, at about 40° BTDC, the piston has completed about 90.5% of its exhaust stroke leaving only 9.5% of its displacement volume, plus the diminutive combustion chamber volume unoccupied.
  • the air charge will have a velocity similar to that of the rising piston and virtually no expansion of the charge will take place before the piston reaches top dead center (TDC).
  • the advancing piston prevents admission of a charge volume greater than the volume of the combustion chamber (whose pressure equilibrates with the manifold- reservoir pressure) at the time of the closing of the intake valve 'i', at about 10 0 BTDC. Combustion will begin before top dead center (BTDC) for the utmost in efficiency. As stated, in this particular arrangement if the compression ratio is 16:1 the expansion ratio will be 48:1.
  • the gases are expanded to three times their initial volume.
  • one stage of compression could be done in the compressor cylinder 5 and the slightly larger volume of charge could be received in the expander cylinders 2, 3 and 4 and a second stage of compression could then be accomplished in the expander cylinders, the compression ratio being established by the volume of the three combustion chamber in relation to the total displaced volume of the single compressor cylinder.
  • the exhaust gases are discharged via an exhaust manifold 21 and the scavenging would be extremely efficient.
  • each piston displaces about 89.4% of its total cylinder volume in the exhaust stroke (displaced volume/total volume).
  • Similar scavenging efficiencies can be realized in the engine according to this invention. For example, if the intake valve 'i' opened at 40 o BTDC and the exhaust valve closed at 40 0 BTDC the stroke of the piston would be 90.54% complete. Therefore, 90.54% of the displacement volume of 522.3 cc (same 4.2 liter engine) is 472.9 cc. This amount divided by the total volume of the cylinder of the engine of this invention is 87.8% of volume displaced (and scavenged).
  • FIG 12 there is shown a similar engine arrangement to that illustrated in Figure 3 in which like parts are designated like reference numerals with the addition of suffix 'b' and in which a projection 150, Figure 12, affixed to the crown of expander cylinder piston 6b, closes the opening of the combustion chamber 151 at somewhere near 40 degrees before top dead center (BTDC) as piston 61) rises in its exhaust stroke.
  • BTDC top dead center
  • This arrangement facilitates exhaust scavenging by allowing the exhaust valve to remain open past TDC and by virtually displacing all of the burned gases while preventing the charge, which is passing the intake valve into the combustion chamber, from entering the cylinder proper.
  • the projection 150 may be fitted with a compression ring 152 residing inside the opening of the combustion chamber as shown in Figure 13.
  • Figure 14 is a diagram for suggested valve timing and can be used with the arrangement shown in Figure 12 for improved scavenging for all of the designs of this invention.
  • the suggested operation is in this manner.
  • the exhaust valve opens near bottom dead center (BDC) and as the piston 6b rises, it expresses the burned gases through the exhaust valve 'e' (not shown) about 40 degrees before top dead center (BTDC), the intake valve opens, at approximately the same time the projection 150 on top of the piston occludes the outlet of the combustion chamber 151 effectively sealing it.
  • the piston has completed 90% of its scavenging, therefore, it only has 10% of further travel.
  • the piston stroke is four inches, then the amount of stroke remaining would be 4/10 inch. Therefore, the projection on the piston would need to be only 4/lOths inch high to seal the combustion opening as the intake valve opens at 40 degrees BTDC. As illustrated in Figure 14, the exhaust valve remains open as much as 30 degrees past TDC.
  • the diagram in Figure 14 illustrates valve timing in which at 40 degrees BTDC the projection 150 on piston 6b closes combustion chamber port 151 and at the same time fresh charge begins to enter intake valve 'i'.
  • the piston continues to rise until there is practically zero clearance with the face of the engine head, expelling virtually all of the exhausted gases.
  • the intake valve is opened, pressure equilibrium is established between the combustion chamber 151 and the manifold 16b.
  • the intake valve closes and fuel is injected and ignited at maximum compression for greatest efficiency.
  • TDC top dead center
  • the exhaust valve 'e' closes.
  • the pressure of the burning gases is expanded against first the piston valve crown 150 and then into the cylinder and against the entire piston crown after the crank angle is 40degrees past top dead center.
  • the charge is expanded against the piston for the full length of the expansion stroke.
  • the compression ratio is established by the total volume of all of the combustion chambers which are supplied by a single compression cylinder, divided into the displaced volume of the single compressor cylinder. For a 2 liter four cylinder engine, this would be 500 cc divided by 31.25 for a compression ratio of 16:1. The combustion chamber volume of this engine would be only 10.4 cc per cylinder or the 31.25 cc for the three firing cylinders.
  • the intake manifold 16 must withstand high pressures this will not add to the weight of the engine because the volume of air charge flowing through it should not be more than 1/16th to 1/8th of the volume passing through the manifold of a conventional engine as the charge is already partially, or preferably, completely compressed. This small volume of charge allows the manifold to have a small inside diameter.
  • the manifold 16 should be small enough for the heavier charge to have sufficient velocity to charge the expander cylinders 2, 3 and 4 but nevertheless should nave enough volume so that there would be no appreciable pressure drop when an expander cylinder is charged.
  • the intake valves 'i' of the engine 1 can be smaller and lighter (requiring lighter springs) and indeed may be shrouded with no loss of volumetric efficiency.
  • Other means besides shrouding for providing a tangential charge direction can also be used.
  • the intake valve will be open for a short time only (such as 30 or 40 0 ), this will be about the 1/8th of the time (or crank angle) that a conventional Otto cycle engine intake valve is normally open. Yet, the volume of charge passing the intake valve, assuming a 16:1 compression ratio, is only 1/48th (one-third of the normal charge already compressed) of the volume passing the intake valve of the Otto cycle engine. In the three or six cylinder engine the volume entering the combustion chamber will be only 1/32 that passing the intake valve of a conventional engine.
  • Fuel may be injected directly into each of the expander cylinders 2, 3 and 4 or into the individual inlet ports.
  • the quantity of fuel may be proportionate to the engine operating conditions by varying the effective stroke of a fuel pump- by varying the opening time of a fuel injection nozzle fed from a constant pressure main or by varying the rate of flow through the injection nozzle.
  • a carburetor may be placed in front on the compressor cylinder 5 and used for maintaining the ratio of fuel to air in the region of the stoichiometric ratio.
  • the engine may be throttled near the atmospheric intake conduit 15 by means of a butterfly valve (not shown) in order to prevent the engine wasting work by having to compress more air than needed to maintain the stoichiometric fuel to air ratio.
  • a means is described later for reducing or eliminating required throttling in the spark ignited version or mode.
  • Figure 2 shows one means of utilizing automatic one-way valves in the compression cylinder 5. While reed type valves 30 (admission), 31 (outlet) are illustrated on the compressor cylinder 5, other valve types, such as sliding valves or sleeve valves could be used.
  • FIGS 3 and 12 of the drawings illustrate one means of operating the intake valves 'i' of the power cylinders of the engine with reference to cylinder 2.
  • the speed of the camshaft 20 is arranged to be the same as that of the crankshaft 10 and is driven from the crankshaft by a gear 22 on the crankshaft and sprocket drive 23 shown in Figure 1.
  • Large cam 24 or 246 operates push-rod 25 or 25b and rockerarm 26 or 26b to activate intake valve 'i' which opens at about 40° BTDC and closes at about 10° BTDC .
  • Figure 4 shows how cam 27 operatespush-rod 28 and rockerarm 29 to activate exhaust valve 'e' which opens at approximately bottom dead center (BTDC) and closes at 40°-35° BTDC in the first design.
  • BTDC bottom dead center
  • the exhaust valve may be held open past top dead center for better scavenging if desired as illustrated in Figures 12 and 14.
  • a bypass line 38 with a one way valve 39 and a blocking valve 40 could be placed in the exhaust manifold 21 in order to direct the pumped air into the manifold 16 for quicker build-up of compression.
  • a second means to facilitate fast starting would be to open a valve leading from a compressed air reservoir to the cylinders. This would supply compressed air for instant firing of the cylinders or could be used to rotate the engine for starting, as described later.
  • the air reservoir could be supplied by an air-compressor retarder brake described with reference to Figure 11 or by any other method.
  • Tremendous swirl can be produced in the combustion chamber by controlling the angle of the inlet port with respect to the cylinder radius or by the use of a shrouded intake valve.
  • the resulting turbulence helps promote combustion by intermixing burned and unburned gases at the flame front as it progresses across the combustion chamber. This feature alone should make NO x and HC emissions negligible and virtually eliminate CO emissions. The extra burning time of the extended expansion process should then further reduce HC emissions to only a trace.
  • FIG. 8 of the drawings there is shown a similar 4-cylinder engine 42, in which like parts are designated like reference numerals with the addition of suffix 'a', and in which additional mid-cylinder exhaust ports 43, 44 and 45 are provided in the walls of the expander cylinders 2a, 3a and Aa respectively, in order to improve the scavenging efficiency.
  • Such ports 43-45 would be uncovered by their associated pistons 6a-8a respectively at the lowest point of the piston stroke. As the exhaust ports 43-45 are uncovered, the pressure in the cylinders could expel much of the exhausted gases to the atmosphere.
  • a step-up gear set 46 can be placed on the crankshaft 10a and geared to drive a scavenging type blower 47 in order to inject fresh air into the ports 43-45 as they are uncovered by their associated pistons 6a-8a, respectively.
  • the associated exhaust valves of each power cylinder 2a-4a would be opened at approximately the same time as the ports 43-45 were uncovered.
  • the exhaust valves are open from before BDC until about 40-45 BTDC and the piston itself displaces (scavenges) 90% of the burnt gases through the exhaust valves. Therefore, if the blower system 46-47 is added, only a small amount of fresh air need be supplied in order to drive some of the burnt gases through the exhaust valve and to dilute the remainder of the gases which are then scavenged by the stroke of the associated piston.
  • the single compressor cylinder could be double acting (not shown) although the basic operation of the engine would remain the same.
  • the compressor cylinder would compress an air charge to a volume sufficient to supply the three power cylinders with one-half to two-thirds of the normal volume of charge depending on the expansion ratio required.
  • a 5-cylinder engine in which one of the cylinders comprised a double acting compressor cylinder would supply four expander (power) cylinders whose combustion chambers are half the volume of a conventional engine. This arrangement will produce four power strokes per revolution with the expansion ratio being twice the compression ratio.
  • any of the 4-cylinder constructions described above could be doubled or alternatively three compressor cylinders could compress the air charge for five power cylinders.
  • the former would produce six power strokes per revolution and the latter would produce five.
  • the combustion chambers could be from 50% to 60% of normal volume according to the expansion ratio desired.
  • the engines may be fueled by means of gasoline, gas or diesel or indeed the engine can be constructed for hybrid operation as a multi-fuel engine.
  • the smaller charge exploded would permit a lighter construction for the compression ignition engine arrangement and will also provide quieter operation for compression ignition (CI) engines.
  • FIG. 9 of the drawings there is shown a schematic transverse sectional view through a six cylinder internal combustion engine having two compressor cylinders 68 and 69 and four expander (power) cylinders 70, 71, 72 and 73 and associated pistons 103, 104, 105, 106, 107 and 108 all connected to a common crankshaft 74 by means of connecting rods 75-80 respectively.
  • the combustion chambers of the expander cylinders are preferably dimensioned to be no more than one-half the volume of that of a conventional engine at a similar compression ratio and therefore the expansion ratio of the engine is at least double that of a conventional engine. For example, at a compression ratio of 16:1 the combustion chamber would be about one-quarter the volume (one-half the normal charge compressed to the higher ratio) of an ordinary engine and the expansion ratio would be 32:1.
  • Each cylinder is a two-stroke cylinder and is scavenged by displacing the burnt gases during the exhaust stroke of the piston. Hence, virtually no air is used in scavenging.
  • the working piston rises displacing the exhaust gases via an exhaust manifold 83, the associated intake valves (109-112) open so that the charge begins to flow at about 40 BTDC and the associated exhaust valves (115-118) close at about 40°BTDC.
  • the enhanced scavenging system illustrated in Figures 12 and 14, and described more fully in the description of the engine of Figure 1, would allow the exhaust valves to remain open past top dead center without allowing the mixing of incoming charge and exhaust gases.
  • the intake valve can have a shroud on one side which directs air charge flow into a very turbulent swirl as previously described. Fuel is injected at the time the intake is in progress or as soon as the intake valve is closed at about 10 BTDC. When the intake valve closes the charge is ignited by spark plug or by means of auto ignition.
  • the volume of the entering air charge in the preferred embodiment is no greater than 1/32nd of that passing through the intake valve of a conventional engine and therefore a good volumetric efficiency is achieved. This gives each of the expander cylinders 70 and 73 one power stroke per revolution so that a total of four power strokes per revolution is produced by the six cylinder engine which, of course, is equal to the number of power strokes of a conventional four-stroke eight-cylinder engine.
  • valves of the power cylinders could be operated as shown in Figures 1, 3 and 6 or in the system illustrated in Figures 12 and 14.
  • the compressor cylinders could be arranged as shown in Figure 2.
  • the manifold 82 would be insulated for compression ignition operation.
  • the air charge could be completely compressed by the compressor cylinders 68 and 69 or, it is also envisaged that the compression could take place partly in the compressor cylinders 68 and 69 and then this charge could be further com- pressed by the expander cylinders 70 to 73.
  • Athree cylinder engine arranged to operate in a similar manner to the six cylinder engine just described is also envisaged.
  • only one compressor cylinder would be provided which would supply a compressed air charge to two expander cylinders thus producing two power strokes per revolution to equal the smoothness of a four-cylinder four-stroke cycle engine.
  • This arrangement would be the same as shown in Figure 1 with one power cylinder removed and the volume of the combustion chambers would ideally be no greater than one-half that of a conventional engine at a similar compression ratio.
  • Either of the two schemes of Figures 4 and 5 or Figures 12 and 14 may be used for scavenging.
  • Reduced throttling can be achieved in any spark ignited engine of this invention which has a plurality of compressor cylinders in the following manner. At any time the atmospheric air intake manifold pressure dropped appreciably below ambient pressure, for example near half throttle, the outlet from one or more of the compressor cylinders could be closed by a shutoff valve. Work done in compressing this captive charge is recovered as the charge expands on the back stroke of the piston with zero net induction pumping done by that cylinder.
  • Throttling may be eliminated completely in spark ignited engines as illustrated in Fig. 1 by providing late fuel injection into the combustion chamber and allowing combustion to begin in the injected spray.
  • the violet swirling motions of the gases will insure that very lean mixtures will burn completely.
  • FIG. 10 of the drawings there is shown a six-cylinder reciprocating internal combustion engine in which all the cylinders 86-91 and associated pistons 119-124 operate on a two-stroke cycle and all cylinders are used for producing power to a common crankshaft 98 via connecting rods 92-97 respectively.
  • This engine is characterized by a more extensive expansion of the burned gases and a greater charge turbulence with combustion beginning at maximum compression.
  • the engine can operate at a higher compression ratio than is usual.
  • the engine is constructed much the same as a four-stroke cycle internal combustion engine but with a number of significant differences.
  • the combustion chamber of each cylinder is preferably made no greater than one-half to one-third the usual size for the compression ratio desired and according to the ' expansion ratio decided upon.
  • the cam shaft (not shown) is geared to turn at the same ⁇ speed as the crankshaft in order to open and close the inlet (125-130) and exhaust (131-136) valves once during each revolution of the crankshaft. Compression takes place in one or more stages before the air charge is admitted to the combustion chambers of the cylinders and the intake manifold becomes a high pressure manifold reservoir.
  • Fuel injectors are used to inject fuel directly into the combustion chambers except for natural gas or propane operation which can be mixed in an EMPCO type carbueretor.
  • An efficient high compression air compressor 99 is placed between the air intake 15 and the working cylinders.
  • any external source of compressed air can replace the compressor 99 and therefore the engine can operate on waste compressed air for further fuel economy.
  • the pressure ratio can be increased at will until the pressure ratio (nominal compression ratio) is equal to or surpasses the expansion ratio for greater power as the load demands. This could be accomplished simply by increasing the speed of the compressor.
  • the compressor 99 aspirates air and compresses it into the manifold-reservoir 100.
  • a check valve at 101 may be used if compressor pressure pulsations are great.
  • the manifold reservoir 100 contains such a volume that there is no appreciable drop in over-all pressure as the cylinders 86-91 are charged sequentially.
  • the working piston ascends to about 40 0 BTDC (see valve timing schemes shown in Figures 5 and 14) which displaces the gases when its travel is almost to the end of its associated cylinder. This expels 90% of the burnt gases through the exhaust valve (into the exhaust manifold 137) which opens as the piston begins its exhaust stroke.
  • the piston is then at about 40o BTDC.
  • the intake valve then opens and an increment of the compressed air charge enters through a valve (can be shrouded) as the piston continues its stroke which is 90% complete. Fuel can be injected at the same time (or as soon as the intake valve is closed.)
  • the high pressure air, the persistency of flow and the small volume of the charge assures a high volumetric efficiency.
  • the intake valve then closes at about 10 0 BTDC and the mixture is ignited. In this manner combustion begins at maximum compression but the air charge has at least two to three times the expansion of an equivalent Otto cycle engine.
  • the expansion ratio will be twice the compression ratio and a one-third normal volume combustion chamber will triple the expansion ratio. If the compression ratio is 16:1, the expansion ratio can be either 32:1 or 48:1, respectively.
  • Enhanced scavenging may be achieved if desired by use of the scavenging system shown in Figures 12 and 14. In this scheme the mouth of the combustion chamber is blocked at about 40° BTDC and the exhaust valve is held open past top dead center, and the intake valve is opened at the time the combustion is blocked. This scheme is better described in the description of the engine of Figure 1.
  • one stage of compression say 8:1 could be done in the compressor 99 and the charge received and further compressed in the expander cylinders.
  • a reciprocating internal combustion engine may have only one compressor cylinder for use in charging a single expander (power) cylinder i.e. a two- cylinder engine.
  • the expander cylinder would be of greater volume than the compressor cylinder.
  • Preignition will not be a problem in the engine of these designs becausa the residence time of the fuel is less than that required for preignition to occur.
  • the power of compression ignition engines operating in this working cycle can be greatly increased by supercharging,
  • the inlet pressure can be boosted from a slight boost up until the theoretical compression ratio equals the expansion rati).
  • Some locomotives operate with a supercharge boost of three atmospheres which, with a compression ratio of 12:1, produces a theoretical compression ratio of 48:1.
  • Some intercooling or aftercooling would likely be required with very high pressure boosts in order to lessen NO x emissions in CI engines.
  • the power of spark ignition engines can be greatly increased by similarly boosting the inlet air pressure.
  • This working cycle may under certain conditions, such as when used in a compression ignition engine at very light loads, result in the combustion gases expanding to pressures less than atmospheric.
  • the nominal compression ratio can be increased until it is equal to the expansion ratio by increasing supercharge boost or by closing off one or more of the expander cylinders. The latter can be done by deactivating their intake and exhaust valves along with their respective fuel injector(s).
  • one expander cylinder could be closed to increase the compression ratio to one-half the expansion ratio. If, under very light loads the pressure at the exhaust valve was still negative, a second expander cylinder could be closed to produce a compression ratio equal to the expansion ratio. With an eight-cylinder engine, one cylinder could be closed at a time for finer control of the compression ratio.
  • the expansion ratio is double the compression ratio.
  • one expander cylinder could be closed to increase the compression ratio to two-thirds the expansion ratio.
  • Two could be closed to produce equal compression and expansion ratios. Aftercooling would not likely be required because now the lightly loaded engine would be using much less fuel and grams NO emissions per mile should not exceed limits.
  • the 1899 Daimler auto engine provided such a means by removing an extra member from between the cam follower and the valve lifter push rod. This allowed the valve spring to hold the valve closed until such time as the spring loaded intermediate member was released.
  • An electronic system of valve control is manufactured by Eaton Corporation and has been used in several automotive engines. This latter system allows the releasing of the rocker arm pivot support in order to deactivate the valve.
  • This system provides electronic controls which can sense exhaust manifold pressure and cut out the necessary number of expander cylinders at such a time the exhaust manifold pressure drops to or below ambient pressure.
  • a pressure sensor, 102 in Figure 9 could be placed in the exhaust manifold and monitored. The fuel rate could then be adjusted so that there would always be a slight positive pressure in the exhaust manifold. This sytem would work well in a constant load, constant speed engine in particular.
  • FIG. 11 of the drawings additional fuel savings can be achieved in the engines described hereinbefore by use of an economizer constructed as an air compressor retarder brake.
  • This six-cylinder engine is similar to the engine shown in Figure 9 in which like parts are designated by like reference numerals with the addition of the suffix 'a'.
  • the air retarder brake illustrated has a compressor 138 operatively connected to the drive shaft of vehicle or geared to the engine and stores energy produced during braking or downhill travel which is utilized to supply compressed air to the engine power cylinders via the transfer manifold of 82a.
  • Such an economizer would be coupled with an air reservoir 139 and during the time in which the economizer reservoir air pressure was sufficiently high for use in the power cylinders of the engine, the engine compressor could be clutchably disengaged so that no compression work would be required of the compressor.
  • a relief valve 140 prevents excess build up of pressure in the air reservoir.
  • One way valve 14 1 - allows air from the reservoir to be transferred to the manifold when the pressure in the reservoir 139 is higher than in the transfer manifold 82a.
  • each compression cylinder of the engine could also be deactivated during this reserve air operation time by shutting off the admission valve so that no net work would be done by the compressor(s) until the manifold- reservoir pressure dropped below operating levels.
  • NMEP net mean effective pressure
  • This economizer or alternatively any other suitable type of air pump may also be used to prevent excessive manifold pressure fluctuation in any of the designs of this invention, if it is found desirable.
  • the engine needs no compression build-up for starting and as soon as the shaft is rotated far enough to open one intake valve the compressed air and fuel would enter and be ignited for "instant" starting.
  • the compressed air could be used to rotate the engine for starting by opening simple valves at the top of the cylinder as is common in large diesel engines, thus eliminating the need for a starter motor.
  • An additional means, to those already suggested, of facilitating cranking of the engine is to hold the intake valve 'i' or the bypass valves 35, 36 and 37 open during the full downstroke of the associated piston thereafter closing the intake valves, holding the exhaust valves closed and then beginning the upstroke of the piston, adding the fuel (if not premixed) and igniting it near the completion of the upstroke, the next downstroke becoming the power stroke.

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Combustion Methods Of Internal-Combustion Engines (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)
EP82300475A 1981-02-02 1982-01-29 Brennkraftmaschine Expired EP0057591B1 (de)

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
US23075281A 1981-02-02 1981-02-02
US230752 1981-02-02
US32792281A 1981-12-08 1981-12-08
US327922 1981-12-08

Publications (3)

Publication Number Publication Date
EP0057591A2 true EP0057591A2 (de) 1982-08-11
EP0057591A3 EP0057591A3 (en) 1983-11-02
EP0057591B1 EP0057591B1 (de) 1989-12-20

Family

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EP82300475A Expired EP0057591B1 (de) 1981-02-02 1982-01-29 Brennkraftmaschine

Country Status (5)

Country Link
EP (1) EP0057591B1 (de)
AU (1) AU7925782A (de)
BR (1) BR8200376A (de)
CA (1) CA1188938A (de)
DE (1) DE3280068D1 (de)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0086348B1 (de) * 1982-02-11 1986-04-30 Robert Bosch Gmbh Verfahren zur Versorgung einer Brennkraftmaschine mit Kraftstoff und Kraftstoffversorgungsanlage zur Durchführung des Verfahrens
WO1993021433A1 (en) * 1992-04-08 1993-10-28 Frederick Arthur Summerlin Internal combustion engine
CN101852090A (zh) * 2010-05-10 2010-10-06 何正品 空气机械动力头

Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1679958A (en) * 1926-03-31 1928-08-07 Automotive Valves Corp Internal-combustion engine
US1955976A (en) * 1932-07-18 1934-04-24 Hans P Rasmussen Internal combustion engine
US2391163A (en) * 1943-06-15 1945-12-18 Sellers E Jessup Art of compressing fluid
US2545793A (en) * 1947-06-23 1951-03-20 Ricardo & Co Engineers Internal-combustion engine operating on the four-stroke cycle with compression ignition
US3408811A (en) * 1967-07-24 1968-11-05 John Donald Wishart Internal combustion engines
US4084556A (en) * 1976-05-14 1978-04-18 Villella Tony R Internal combustion engine
US4202300A (en) * 1978-02-22 1980-05-13 Frank Skay Internal combustion engine
US4210109A (en) * 1976-12-02 1980-07-01 Nissan Motor Company, Limited Multi-cylinder internal combustion engine
GB2071210A (en) * 1980-02-29 1981-09-16 Kaltenegger B Four-stroke engine with a charging piston pump

Patent Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1679958A (en) * 1926-03-31 1928-08-07 Automotive Valves Corp Internal-combustion engine
US1955976A (en) * 1932-07-18 1934-04-24 Hans P Rasmussen Internal combustion engine
US2391163A (en) * 1943-06-15 1945-12-18 Sellers E Jessup Art of compressing fluid
US2545793A (en) * 1947-06-23 1951-03-20 Ricardo & Co Engineers Internal-combustion engine operating on the four-stroke cycle with compression ignition
US3408811A (en) * 1967-07-24 1968-11-05 John Donald Wishart Internal combustion engines
US4084556A (en) * 1976-05-14 1978-04-18 Villella Tony R Internal combustion engine
US4210109A (en) * 1976-12-02 1980-07-01 Nissan Motor Company, Limited Multi-cylinder internal combustion engine
US4202300A (en) * 1978-02-22 1980-05-13 Frank Skay Internal combustion engine
GB2071210A (en) * 1980-02-29 1981-09-16 Kaltenegger B Four-stroke engine with a charging piston pump

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0086348B1 (de) * 1982-02-11 1986-04-30 Robert Bosch Gmbh Verfahren zur Versorgung einer Brennkraftmaschine mit Kraftstoff und Kraftstoffversorgungsanlage zur Durchführung des Verfahrens
WO1993021433A1 (en) * 1992-04-08 1993-10-28 Frederick Arthur Summerlin Internal combustion engine
CN101852090A (zh) * 2010-05-10 2010-10-06 何正品 空气机械动力头

Also Published As

Publication number Publication date
AU7925782A (en) 1982-08-12
EP0057591B1 (de) 1989-12-20
DE3280068D1 (de) 1990-01-25
CA1188938A (en) 1985-06-18
EP0057591A3 (en) 1983-11-02
BR8200376A (pt) 1982-11-23

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