WO2019082403A1 - Internal combustion engine - Google Patents

Internal combustion engine

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Publication number
WO2019082403A1
WO2019082403A1 PCT/JP2018/000151 JP2018000151W WO2019082403A1 WO 2019082403 A1 WO2019082403 A1 WO 2019082403A1 JP 2018000151 W JP2018000151 W JP 2018000151W WO 2019082403 A1 WO2019082403 A1 WO 2019082403A1
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WO
WIPO (PCT)
Prior art keywords
internal combustion
combustion engine
intake
fuel
valve
Prior art date
Application number
PCT/JP2018/000151
Other languages
French (fr)
Japanese (ja)
Inventor
正裕 井尻
Original Assignee
正裕 井尻
Priority date (The priority date 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 date listed.)
Filing date
Publication date
Application filed by 正裕 井尻 filed Critical 正裕 井尻
Publication of WO2019082403A1 publication Critical patent/WO2019082403A1/en

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Classifications

    • 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
    • F02B25/00—Engines characterised by using fresh charge for scavenging cylinders
    • F02B25/14—Engines characterised by using fresh charge for scavenging cylinders using reverse-flow scavenging, e.g. with both outlet and inlet ports arranged near bottom of piston stroke
    • F02B25/16—Engines characterised by using fresh charge for scavenging cylinders using reverse-flow scavenging, e.g. with both outlet and inlet ports arranged near bottom of piston stroke the charge flowing upward essentially along cylinder wall opposite the inlet 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
    • F02B31/00—Modifying induction systems for imparting a rotation to the charge in the cylinder
    • F02B31/04—Modifying induction systems for imparting a rotation to the charge in the cylinder by means within the induction channel, e.g. deflectors
    • F02B31/06—Movable means, e.g. butterfly valves
    • 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
    • F02B33/20—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps with pumping-cylinder axis arranged at an angle to working-cylinder axis, e.g. at an angle of 90 degrees
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D19/00—Controlling engines characterised by their use of non-liquid fuels, pluralities of fuels, or non-fuel substances added to the combustible mixtures
    • F02D19/02—Controlling engines characterised by their use of non-liquid fuels, pluralities of fuels, or non-fuel substances added to the combustible mixtures peculiar to engines working with gaseous fuels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D19/00—Controlling engines characterised by their use of non-liquid fuels, pluralities of fuels, or non-fuel substances added to the combustible mixtures
    • F02D19/06—Controlling engines characterised by their use of non-liquid fuels, pluralities of fuels, or non-fuel substances added to the combustible mixtures peculiar to engines working with pluralities of fuels, e.g. alternatively with light and heavy fuel oil, other than engines indifferent to the fuel consumed
    • F02D19/08—Controlling engines characterised by their use of non-liquid fuels, pluralities of fuels, or non-fuel substances added to the combustible mixtures peculiar to engines working with pluralities of fuels, e.g. alternatively with light and heavy fuel oil, other than engines indifferent to the fuel consumed simultaneously using pluralities of fuels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D41/00—Electrical control of supply of combustible mixture or its constituents
    • F02D41/02—Circuit arrangements for generating control signals
    • F02D41/04—Introducing corrections for particular operating conditions
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D45/00—Electrical control not provided for in groups F02D41/00 - F02D43/00
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/30—Use of alternative fuels, e.g. biofuels

Definitions

  • the present invention relates to the combustion means of a two-stroke internal combustion engine, and to the combustion means and valve drive mechanism of a four-stroke internal combustion engine.
  • the two-stroke internal combustion engine (two-stroke one-stroke engine) generates twice as much combustion stroke as the four-stroke internal combustion engine per revolution, and thus a large output can be obtained.
  • the exhaust is added between the combustion stroke and the compression stroke. Since the scavenging that performs gas exchange with compressed intake is performed, the strokes of the combustion stroke and the compression stroke are reduced, so the expansion stroke decreases in the combustion stroke, the output efficiency decreases by the compression stroke, and the compression ratio by the compression stroke decreases in the compression stroke pressure.
  • the capacity of the supercharger may need to be increased more than necessary, and there is a problem that the supercharger becomes larger.
  • an accumulator chamber larger than a piston stroke volume of at least one cylinder is provided in an intake passage between the turbocharger and the combustion chamber, A check valve can be arranged on the upstream side to equalize the required supercharging amount during one rotation of the engine, and the capacity of the supercharger can be reduced by that amount, which can contribute to downsizing
  • Patent Document 1 equipped with a motor.
  • a compression-ignition "split-cycle” engine is a compression-ignition engine that emits less dust and releases nitrogen oxides, and there is also a split-cycle engine (Patent Document 2) that can be supercharged by a turbosupercharger (claim 12) . Since these superchargers (patent documents 1 and 2) require a pump capacity of a capacity larger than the cylinder displacement, there is a problem that further miniaturization of the internal combustion engine becomes difficult.
  • a spherical combustion chamber In an internal combustion engine, in order to shift to a combustion stroke without attenuating an intake swirl (hereinafter referred to as "swirl"), which is an intake flow, for the purpose of improving a combustion speed (flame propagation speed), a spherical combustion chamber
  • swirl an intake swirl
  • the operation of the cam mechanism driven by the engine is transmitted to the valve by the mechanical transmission mechanism to open and close the valve, so the shape of the combustion chamber is restricted to the pentroof shape, causing swirl.
  • the design is constrained by the undesirable shape.
  • the mechanical transmission mechanism a hydraulic transmission mechanism, a spherical combustion chamber and a liquid pressure drive valve radially disposed in the combustion chamber
  • an internal combustion engine Patent Document 3 including a cam mechanism driven by an engine and a fluid mechanism that transmits the operation of the cam mechanism to the valve via fluid pressure to open and close the valve. Since the combustion chamber is spherical, there is an effect of suppressing the attenuation of the swirl, and a reduction in the surface area of the combustion chamber and the cooling loss.
  • a positive displacement pump driven by a four-cycle internal combustion engine is composed of a rotor, a tubular cam, and vanes or plungers provided on the rotary shaft of the rotor.
  • a valve drive mechanism (patent document 4) of an internal combustion engine which can arrange the hydraulic circuit of the above and has a high reliability of the hydraulic pressure supply means by a simple structure, and can be made compact and inexpensive.
  • the combustion range (Vol%) of hydrogen (4.1 to 75) is wider than gasoline (1 to 7.8), and the minimum ignition energy (mj) of hydrogen (0.02) is gasoline (0%) for combustion of internal combustion engines. .24) smaller and hydrogen (346) is larger than gasoline (42) because the maximum burning rate (cm / s) is larger than hydrogen (42), so it is easy to ignite and has a large blast pressure. There is a problem of low energy density.
  • a gas different from the main fuel for forming the fuel mixture for the purpose of expanding the lean limit of the fuel mixture by effectively using gaseous fuel such as hydrogen and realizing the improvement of the thermal efficiency and the improvement of the exhaust A method of gaseous fuel addition of an internal combustion engine, wherein fuel is added into a cylinder, a first addition step of adding the gaseous fuel during an intake stroke, and a second addition step of adding the gaseous fuel during a compression stroke And a gaseous fuel addition method for an internal combustion engine (Patent Document 5).
  • An internal combustion engine comprising an in-cylinder injection nozzle for injecting gasoline fuel into the cylinder and a port injection nozzle for injecting hydrogen fuel at an intake port for the purpose of effectively suppressing the occurrence of knocking;
  • a pair of intake ports and a pair of intake valves are provided, control means for controlling the opening and closing of the intake valves is provided, and the port injection nozzle performs intake
  • the hydrogen fuel is injected at the port, and one of the intake valves of the pair of intake valves provided in one of the cylinders is opened and the other of the intake valves is closed to perform intake of air in the cylinder.
  • An internal combustion engine (patent document 6) which generates a swirl.
  • the ignition point is higher than hydrogen (500 ° C) than diesel fuel oil (225 ° C), so an HCCI engine (a diesel engine etc. that pre-mixes hydrogen by controlling the temperature at the time of adiabatic compression ignition)
  • the flammability can be improved by using a premixed compression auto-ignition) or an SPCCI engine (compression-ignition combustion by spark control).
  • a control device for a homogeneous charge compression ignition combustion type internal combustion engine capable of stably operating over a wide range of operation range, a fuel supply system for supplying light oil or a mixed fuel containing light oil to the engine, and a gas supply system for supplying hydrogen to the engine
  • a required premixed gas operation unit that has, as data, a plurality of combustion waveforms that change depending on the concentration of hydrogen addition, and the thermal efficiency is increased according to the state of the engine, from among the plurality of combustion waveforms
  • a control device for an internal combustion engine capable of reducing the amount of PM and NOx produced and improving the thermal efficiency of the engine by selecting an appropriate one and determining the hydrogen addition concentration supplied to the engine so as to match the combustion waveform.
  • multiple intake ports communicating with the cylinder, a hydrogen injector that adds hydrogen-containing gas to the intake, and light oil fuel
  • a plurality of intake ports include a secondary intake port which is a helical port and a primary intake port which is a tangential port, and the hydrogen injector is configured to set a secondary intake port among the intake ports.
  • Patent Document 8 There is a diesel internal combustion engine (Patent Document 8) in which hydrogen-containing gas is added only to the intake air introduced into the cylinder via.
  • Fuel hydrogen can be generated by fuel reforming, there is a problem of consuming fuel.
  • Hydrogen can also electrolyze water (electrolyte solution) to generate hydrogen and oxygen, propagate ultrasonic vibration to the electrolyte, and efficiently generate hydrogen and oxygen or oxygen from the electrode.
  • Patent Document 9 can be mounted on a hybrid vehicle or the like using hydrogen or hydrogen oxide as a fuel.
  • a small amount of fuel is distributed in the center of the combustion chamber to improve ignition performance, and sufficiently mixed with the surrounding intake air to improve combustion efficiency, or hydrogen has good ignition performance and a large blast pressure.
  • fuel having good combustibility from the center of the combustion chamber where combustion starts It is desirable to arrange the U.S. in layers and mix them uniformly with the fuel to make effective use of the oxygen in the intake.
  • the combustion chamber is formed into a spherical or substantially conical axisymmetric shape, and the strong swirl generated from the two intake ports mixes the intake and fuel, and air such as hydrogen during the compression stroke
  • a four-cycle internal combustion engine that collects lower density fuel by centrifugal separation near the cylinder axis of the combustion chamber in which the spark plug or injector is disposed and separates the remainder into layers according to the concentration of fuel to improve flammability through stratified combustion ( Patent Document 10).
  • the centrifugal separation action by swirl acts between two strokes in the intake stroke and the compression stroke, but in the two-stroke internal combustion engine, one stroke of the scavenging stroke between the exhaust stroke and the compression stroke and the compression stroke.
  • the centrifugal action does not work sufficiently because the stroke is less than half and the stroke is about half.
  • the scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the displacement, the reduction of the filling efficiency due to the scavenging blowout is eliminated, and the initially inflowing intake air whose speed is attenuated is discharged. Since the amount of scavenging air flowing during the stroke is increased, the strong swirl with increased intake air inflow speed improves the combustibility of the internal combustion engine, and further downsizing can be performed by the power increase effect by supercharging.
  • the combustion chamber has a substantially spherical shape or a substantially conical shape, and an intake valve and an exhaust valve are radially arranged in the combustion chamber, an ignition plug or an injector is provided near the intersection with the cylinder axis of the combustion chamber, and an intake port is in the cylinder.
  • the tangential port generates swirl, and the scavenging air supply means actively blows through scavenging to generate a strong swirl in the combustion chamber, and the shape of the combustion chamber smoothly causes the swirl to reach TDC in the compression stroke.
  • the diameter of the swirl is reduced due to the deformation of the shape of the combustion chamber into a dome due to the compression stroke, and the centrifugal action of the strong swirl generated at the center of the cylinder causes a density gradient to be less dense fuel such as hydrogen.
  • Can be separated into appropriate layers, and hydrogen and the like with good combustibility can be collected around the cylinder axis of the combustion chamber to improve the combustibility by stratified combustion, and further spark-ignition internal combustion
  • supplying fuel Seki or diesel engine can be improved combustibility. Since the cylinder head is provided with the intake valve and the exhaust valve, there is no need to mix lubricating oil with fuel, and the problem of soot being generated in the exhaust which is a problem of the two-stroke internal combustion engine is solved.
  • scavenging gas supply means a compressor driven by the internal combustion engine and a scavenging gas amplification means having a simple configuration provided with an air flow rate amplifier in the intake passage can be provided to supply intake air larger than the displacement of the internal combustion engine.
  • the blow-by portion is replenished to improve the combustibility by reliable gas exchange, and after the exhaust valve is closed, pressurized scavenging gas can be supplied to perform supercharging.
  • a plurality of intake valves and exhaust valves are alternately arranged, and two camshafts rotationally driven synchronously by the crankshafts are provided, and the exhaust valves are opened and closed on the two camshafts.
  • One cam can be provided, and the intake valve can be a check valve, or a drive mechanism of the intake valve by a hydraulic unit operated by the cam can be provided, and the combustibility can be improved by the swirl generated in the cylinder axis symmetry.
  • an oxyhydrogen generator operated by an electrical means is provided, and the oxyhydrogen generated by the oxyhydrogen generator, or hydrogen and oxygen are supplied to the internal combustion engine, and storage is easier than hydrogen. It can be used as electrolytic solution storage and energy regeneration means using electrical means.
  • the combustion chamber of the four-stroke internal combustion engine is substantially spherical or substantially conical, and a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber to generate swirls having the same rotational direction.
  • an internal combustion engine provided with a plurality of tangential ports for supplying hydrogen or fuel such as methane having a density smaller than air to an intake system and / or a combustion chamber
  • the rotation of the crankshaft of the reciprocating engine Two parallel camshafts driven at half the number of revolutions of the number are provided, and the two camshafts are provided with cams for opening and closing the exhaust valve, and are opened and closed by hydraulic means interlocking with the cams or
  • the intake valve can be opened and closed by hydraulic means that operates the intake valve with a cam different from the cam, and the combustibility can be improved by the swirl generated in a cylinder axial symmetry.
  • the fuel can not be collected efficiently and reliably near the spark plug in the combustion chamber, and the combustion efficiency can not be improved because the combustible layer which makes the flame propagation uniform in the circumferential direction can not be formed. There is.
  • the first to third aspects relate to the combustion means of the two-stroke internal combustion engine, and the fourth aspect relates to the combustion means of the four-stroke internal combustion engine.
  • a first aspect of the present invention is a two-stroke internal combustion engine provided with an intake valve and an exhaust valve in a cylinder head, comprising: scavenging air supply means capable of supplying scavenging air having a capacity larger than the exhaust amount; the scavenging air supply means driven by the internal combustion engine And a scavenging amplification means provided in the intake passage, the scavenging amplification means comprising a check valve and an air flow amplifier provided downstream of the check valve, the drive flow path of the air flow amplifier being compressed Furthermore, the combustion chamber has a substantially spherical or substantially conical shape, and an intake valve and an exhaust valve are radially arranged in the combustion chamber, and an ignition plug or an injector is connected to the cylinder axis of the combustion chamber.
  • the intake port is a tangential port that generates swirl in the cylinder, and hydrogen, a fuel such as methane having a density smaller than that of air, or a fuel having a density smaller than that of air
  • a fuel such as methane having a density smaller than that of air
  • a fuel having a density smaller than that of air Fuel flowers ignition type internal combustion engine or diesel engine is supplied to the intake system and / or the combustion chamber, a two-cycle internal combustion engine for controlling the supply of the fuel in accordance with the operating condition of the internal combustion engine.
  • a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber, and two parallel camshafts interlocked at the same rotational speed as the rotational speed of the crankshaft are provided, and the exhaust valve Are opened and closed by a cam provided on each of the two camshafts, and the intake valve serves as a check valve, or the intake valve is opened and closed by hydraulic means interlocking with the cam, or 2.
  • the two-stroke internal combustion engine according to claim 1 wherein the two-stroke internal combustion engine is opened and closed by hydraulic means interlocked with another cam.
  • the internal combustion engine is provided with an oxyhydrogen generator operated by electrical means, and hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air. It is a two-stroke internal combustion engine according to claim 2.
  • the combustion chamber is substantially spherical or substantially conical, a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber, and an ignition plug or injector is disposed in the combustion chamber.
  • the intake port is a tangential port that generates a swirl in the cylinder, and fuel such as hydrogen and methane has a density smaller than that of air, or fuel whose density is smaller than that of air and spark ignition
  • Fuel of the internal combustion engine or diesel engine is supplied to the intake system and / or the combustion chamber, and the supply of the fuel is controlled according to the operating condition of the internal combustion engine, and Hydraulic means provided with two parallel camshafts interlocking at rotational speed, opening and closing the exhaust valve by cams provided at the two camshafts, and interlocking the intake valve with the cam It is a four-stroke internal combustion engine that opens and closes or opens and closes the intake valve by hydraulic means linked to a cam other than the cam.
  • the intake valve and the exhaust valve are provided in the cylinder head and the crankcase is not used as the pump chamber, it is not necessary to mix lubricating oil into the intake, and the problem of generating soot in the exhaust can be solved.
  • the combustion chamber has a substantially spherical or substantially conical shape, the intake valve and the exhaust valve are radially arranged in the combustion chamber, the spark plug or the injector is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is in the cylinder. It is a tangential port that generates swirl, and hydrogen and other fuels with a density smaller than air such as methane are supplied to the air system and / or the combustion chamber, so that the hydrogen etc.
  • the intake swirl generated in the combustion chamber is generated by the intake swirl generated in the combustion chamber.
  • the fuel whose density is smaller than that of the air is moved to the cylinder shaft side provided with the spark plug or injector by centrifugal separation, and as shown in FIGS. 3 and 4, the shape of the combustion chamber macroscopically Since it changes from a substantially cylindrical shape to a substantially spherical dome shape, the laminar distribution of fuel in the radial direction of intake moves from the peripheral portion to the central portion, and the swirling diameter of the swirl holds kinetic energy.
  • the scavenging air flow is discharged as a short circuit scavenging air due to the blowout to make a strong swirl, and the strong swirl can be continued without the rotational movement of the swirl being inhibited by the shape of the combustion chamber until the end of the compression stroke.
  • the fuel such as hydrogen is separated into layers as shown in FIG. 5 according to the density gradient of the gas, forming a high concentration combustible layer in the center, and the cylinder axis of the combustion chamber Collecting at the ignition part of the spark plug or the injector provided near the intersection of As shown in FIG.
  • a high concentration layer of hydrogen or the like having a small density (molecular weight) is formed on the cylinder axis side by the centrifugal separation action, and a combustible layer of the main fuel having a density larger than hydrogen is formed around it.
  • the rotation speed can be increased and the output can be increased.
  • the high concentration combustible layer such as hydrogen can be ignited or ignited with certainty, flame propagation propagates uniformly in the circumferential direction sequentially from the high concentration layer side to the low concentration layer, and in the outermost ultra low concentration layer the combustion temperature is Since the temperature rise of the combustion chamber wall surface by flame propagation is suppressed since it is low, the cooling loss is suppressed, and the thermal efficiency of the internal combustion engine is increased by the synergistic effect with the effect of improving the combustibility by the stratified combustion.
  • the knocking phenomenon in which the remaining unburned air-fuel mixture burns rapidly before reaching the flame propagation has a knocking suppression effect due to the stratified combustion (ultra-low concentration layer in the outermost layer) according to the density gradient.
  • scavenging amplification means comprising an air flow amplifier having a compressed air generated by the compressor as a drive flow has an effect of being able to supply scavenging air exceeding the displacement of the internal combustion engine with a simple and small-capacity compressor.
  • the scavenging air amplification means is provided with a check valve on the downstream side of the intake passage, and a reverse flow amplification phenomenon of the air flow amplifier (a flow is reversed due to the pressure increase on the downstream side and a backflow phenomenon with flow amplification on the upstream side)
  • the high pressure drive flow directly flows into the cylinder by the check valve to generate the supercharging effect.
  • the blow through has a cooling effect on the combustion chamber of a two-stroke internal combustion engine having a large heat load. Since fuel can be supplied to the drive flow passage in the latter half of the scavenging timing and the fuel can be uniformly premixed in the intake air by the drive flow simultaneously with supercharging, oxygen in the intake air can be efficiently used for combustion. As shown in FIG. 20, the HCCI engine (premixed compression auto-ignition) or SPCCI engine (compression-ignition combustion by spark control) with the temperature control of the combustion chamber has the effect of improving the combustibility and the output. .
  • a plurality of intake valves and exhaust valves are alternately arranged, and two camshafts rotationally driven synchronously by the crankshafts are provided, and the exhaust valves are opened and closed on the two camshafts.
  • the intake valve is provided as a check valve, or a drive mechanism for the intake valve is provided by hydraulic means operated by the cam, and the combustibility can be improved by the swirl generated in the cylinder axis symmetry, and the valve
  • the internal combustion engine is provided with an oxyhydrogen generator operated by electrical means, and hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air.
  • hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air.
  • the combustion chamber of the four-stroke internal combustion engine is substantially spherical or substantially conical, and a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber.
  • a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber.
  • each cam provided with one cam for opening and closing the exhaust valve, and the intake by hydraulic means operating the intake valve with the cam Since the drive mechanism of the valve is provided, the strong swirl generated by the tangential port arranged opposite to the cylinder axis can improve the combustibility, and the exhaust valve acting on the exhaust pressure at the time of valve opening is cam driven, high speed and high load Reliability during driving is improved.
  • FIG. 7 is a plan view and a peripheral circuit diagram of an internal combustion engine using a hydrogen and a gasoline as fuel, in which tangential ports are provided in a substantially spherical combustion chamber according to Embodiment 2 (corresponding to claim 1).
  • scavenging behavior swirl
  • exhaustion of the compression stroke of the combustion chamber of the said Example 2 FIG. 7
  • FIG. 2 is a characteristic diagram of fuel and air of the internal combustion engine etc. of the second embodiment (FIG. 2), wherein the fuel shows the combustion range and the density, and the air shows the composition ratio and the density.
  • FIG. 2 is a characteristic diagram of fuel and air of the internal combustion engine etc. of the second embodiment (FIG. 2), wherein the fuel shows the combustion range and the density, and the air shows the composition ratio and the density.
  • FIG. 14 is a plan view of a three-cylinder internal combustion engine showing the arrangement of a reciprocating compressor, intake and exhaust valves and passages according to a third embodiment (corresponding to claim 1), and peripheral circuits such as scavenging air amplification means and positive displacement hydraulic pressure supply means.
  • FIG. 8 is a cross-sectional view of the internal combustion engine provided with cooling means for the intake valve and the exhaust valve of the JJ cross section of the third embodiment (FIG. 7). It is explanatory drawing of the construction concept of the 2 cycle internal combustion engine provided with the scavenging-gas amplification means which consists of a non-return valve and the air flow amplifier of Example 4 (Claim 1 claim), and a reciprocating compressor whose phase advanced from the crankshaft. .
  • FIG. 14 is an operation explanatory view of scavenging amplification means of the exhaust stroke initial (S1), scavenging stroke (S2), and compression stroke (S3) of the internal combustion engine of the fourth embodiment (FIG. 9). It is a timing diagram of the internal combustion engine of the said Example 4 (FIG. 9).
  • FIG. 15 It is sectional drawing of the scavenging air amplification means comprised with the variable nozzle type transvector and lift check valve of the prior art of the said Example 5 (FIG. 15).
  • They are the in-cylinder pressure by trial calculation and the timing chart of each cylinder of the internal combustion engine of the said Example 6 (FIG. 17).
  • FIG. 20 is a plan view (1) and an MM cross-sectional view (2) of a configuration explanatory view of a three-cylinder internal combustion engine provided with the scavenging air amplification means of the seventh embodiment (corresponding to claim 1) and a reciprocating compressor.
  • FIG. 18 is an explanatory view of a configuration concept of a two-stroke internal combustion engine in which a reciprocating compressor according to an eighth embodiment (corresponding to claim 1) and a control valve are provided to a scavenging air amplification means.
  • FIG. 21 is a PV diagram based on a trial calculation at the time of flow rate amplification suppression of scavenging air by a control valve provided in the scavenging air amplification means of the eighth embodiment (FIG. 20).
  • FIG. 2 is an explanatory view of a configuration concept of a two-cylinder two-stroke internal combustion engine provided with a scavenging air amplification means as a valve.
  • FIG. 24 is a plan view and a peripheral circuit diagram of a two-cylinder two-stroke internal combustion engine in which an exhaust valve is cam-driven and an intake valve is hydraulically driven via hydraulic means by two camshafts of Embodiment 10 (corresponding to claim 2).
  • FIG. 24 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the tenth embodiment (FIG. 23). It is explanatory drawing of the structural concept of the internal combustion engine which uses hydrogen as a fuel provided with the scavenging air amplification means and the oxyhydrogen generator of Example 11 (corresponding to claim 3).
  • FIG. 16 is an explanatory diagram of a conceptual configuration of an oxyhydrogen generator for adding an ultrasonic wave to a prior art electrolyte solution of Example 12 (corresponding to claim 3), which is an example of the oxyhydrogen generator of Example 11 (FIG. 2). It is.
  • FIG. 26 is an explanatory diagram of a configuration concept of a control system which can be operated as an HCCI engine or an SPCCI engine in the internal combustion engine of the hybrid vehicle in the thirteenth embodiment (FIG. 26).
  • FIG. 33 is a control flowchart for operating the control system of the internal combustion engine of the hybrid vehicle incorporating the internal combustion engine 1t of the thirteenth embodiment as an HCCI engine or an SPCCI engine.
  • FIG. 31 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and combustion of an HCCI engine at the time of fuel injection in the fourteenth embodiment (FIG. 30).
  • Embodiment 15 A four-stroke internal combustion engine which opens and closes an exhaust valve by means of cams provided on two camshafts, and opens and closes an intake valve through hydraulic means by means of a cam different from the cam.
  • FIG. 32 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the fifteenth embodiment (FIG. 32).
  • the internal combustion engine may be a spark ignition internal combustion engine or a diesel engine
  • the fuel supply may be an intake system and / or a combustion chamber
  • the hydraulic pump of the positive displacement hydraulic supply means may be a vane pump It may be a plunger pump.
  • FIG. 1 shows the configuration of a two-stroke internal combustion engine provided with hydrogen fuel by providing an intake valve and an exhaust valve radially disposed in a substantially spherical combustion chamber and a tangential intake port according to Embodiment 1 (corresponding to claim 1). It is explanatory drawing of a concept.
  • FIG. 1 shows a two-stroke internal combustion engine 1 in which an intake valve 46 and an exhaust valve 47 are provided in a cylinder head, provided with scavenging air supply means capable of supplying scavenging air of a larger capacity than the exhaust gas.
  • the scavenging amplification means 5 is provided between the compressor 25 driven at the same time and the intake inflow passage 22 and the intake and outflow passage 23 which are intake passages, and the scavenging amplification means 5 comprises a check valve 55 and a check valve 55.
  • the air flow amplifier 50 is provided downstream, and the drive flow passage 58 of the air flow amplifier 50 is in communication with the discharge port of the compressor 25, and the combustion chamber has a substantially spherical shape with a radius SR.
  • the intake valve 46 and the exhaust valve 47 are arranged radially, the spark plug 11 is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is a tangential port 230 that generates swirl in the cylinder.
  • the fuel density than air is small is supplied to the combustion chamber as a 2-cycle internal combustion engine 1 for controlling the supply of the fuel in accordance with the operating condition of the internal combustion engine 1.
  • the internal combustion engine 1 is a split-cycle two-stroke internal combustion engine 1 provided with a reciprocating compressor 25 having a discharge amount equal to or greater than the displacement of the internal combustion engine 1, and the reciprocating compression is performed on the crankshaft having the same phase as the connecting rod 43 of the output means 4.
  • the connecting rod 253 of the machine 25 is provided, and the diameter ( ⁇ C) of the cylinder 251 of the reciprocating compressor 25 is made smaller than the diameter ( ⁇ E) of the cylinder 41 of the output means 4.
  • the scavenging air of the above capacity is supplied, and the hydrogen stored in the fuel tank 75 is pressurized by the high pressure fuel pump 13 and is supplied to the combustion chamber by the injector 12 as needed.
  • the intake valve 46 is a check valve that does not require a valve drive mechanism
  • the exhaust valve 47 is a hydraulic pressure generated by the positive hydraulic pressure supply means 8 supplied from the hydraulic pressure passage 88 with a hydraulic actuator consisting of a valve cylinder 471 and a valve piston 472 Act to open and close the valve.
  • the positive displacement hydraulic pressure supply means 8 includes a hydraulic pump consisting of a rotor 82 rotating in synchronization with the crankshaft 44, a cam 81 and a vane 83, and a vane 83 of the rotor 82 is added to a third embodiment (FIG.
  • the valve drive mechanism of an internal combustion engine according to the prior art Patent Document 4 that can handle multiple cylinders as in the above.
  • the operation of the internal combustion engine 1 of FIG. 1 is that the two-stroke internal combustion engine 1 consisting of a reciprocating compressor 25 performing a division cycle and an output unit 4 performs scavenging of scavenging air over the displacement of the internal combustion engine 1 by the reciprocating compressor 25; Scavenger inflow speed is increased by supplying swirl to the cylinder 41 from the sensial port 230 to generate swirl and increasing the scavenging amount supplied in the scavenging stroke, and the speed of swirl decreases due to the collision with the exhaust.
  • the air flow is released as a short circuit scavenging air due to the blow through to make a strong swirl, and the strong swirl can be continued without the rotational movement of the swirl being inhibited by the shape of the combustion chamber until the end of the compression stroke.
  • the internal combustion engine 1 is a two-stroke internal combustion engine provided with an intake valve and an exhaust valve at the cylinder head, the scavenging efficiency is lower than that of the uniflow scavenging system with high scavenging efficiency.
  • the combustion efficiency can be improved by improving the
  • the sum of the working angles of the vanes 83 of the displacement type hydraulic pressure supply means 8 and the working angles ⁇ c of the cam profile 881 of the cam 81 indicates the opening timing angle of the exhaust valve 47 by the hydraulic pressure. It becomes.
  • the check valve 875 of the hydraulic pressure auxiliary means 87 communicating with the hydraulic pressure hydraulic passage 88 for opening the exhaust valve 47 performs oil replenishment when hydraulic oil leaks, and thus has a lash adjuster function of the exhaust valve 47.
  • the valve drive mechanism of the exhaust valve 47 may be another drive mechanism as long as the valve arrangement is possible.
  • the intake valve 46 is a check valve, but can be opened and closed by providing a valve drive mechanism.
  • the combustion of hydrogen generates water, and a small amount of water or steam has the effect of assisting combustion, so the exhaust properties are improved, and when it is steam, it expands about 1700 times and increases the pressure in the cylinder of the internal combustion engine There is an output increase effect.
  • the centrifugal separation action of the fuel is as shown in FIG. 2 and FIG. 3 in the behavior and action of intake by making the combustion chamber substantially spherical and making the intake port the tangential port 230 in Example 2 described later.
  • the distribution of the fuel concentration layers formed by the centrifugal separation action of the swirl and the flame propagation at the time of ignition in FIG. 5 are the characteristics of the fuel and air (intake air) that are the basis for the improvement of the centrifugal separation action and combustibility.
  • the operation and configuration of the scavenging amplification means 5 will be described with reference to FIGS. 6 to 6 in FIGS.
  • FIG. 2 is a plan view and a peripheral circuit diagram of an internal combustion engine using a hydrogen and a gasoline as fuel, in which a tangential port is provided in a substantially spherical combustion chamber according to a second embodiment (corresponding to claim 1).
  • FIG. 2 shows scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the exhaust gas amount in a two-stroke internal combustion engine 1g provided with an intake valve 46g and an exhaust valve 47g in a cylinder head,
  • the scavenging air supply means includes a compressor 25g driven by the internal combustion engine 1g, and a scavenging air amplification means 5g between an intake air inflow passage 22g and an intake air outflow passage 23g, which are intake air passages.
  • the intake valve 46g and the exhaust valve 47g are radially disposed in the combustion chamber, the spark plug 11g is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is swirled in the cylinder.
  • tangential port to be used to supply fuel which has a density smaller than that of air such as hydrogen, and gasoline, which is the fuel of a spark ignition type internal combustion engine, to the intake system and the combustion chamber
  • the hydrogen to be pressurized and stored in the fuel tank 75g of the fluid supply means 7g is properly injected with an appropriate amount by the injector 12g provided in the intake and outlet passage 23g, and premixed with the intake air which does not blow when scavenging. Inject at the timing shown in.
  • the gasoline stored in the fuel tank 75g2 of the fluid supply means 7g2 is appropriately injected in an appropriate amount into the combustion chamber by the injector 12g2.
  • the intake valve 46g is a check valve, and the exhaust valve 47g operates the valve cylinder 471g with hydraulic pressure from a positive displacement hydraulic pressure supply means (not shown) supplied from the hydraulic pressure passage 88g to open and close the valve.
  • the configuration and operation of the reciprocating compressor 25g are the same as in the first embodiment.
  • FIG. 3 is an explanatory view of the scavenging behavior (swirl) during scavenging (P1) after exhaust of the compression stroke and after scavenging of the compression stroke (P2) in the combustion chamber of the second embodiment (FIG. 2). is there.
  • the exhaust stroke ends with the closing of the exhaust valve 47g, and the intake flow flowing into the cylinder 41g from the two tangential ports 230g generates two swirls in the cylinder.
  • FIG. 3 one of the two swirls is illustrated.
  • the one swirl is axially compressed by the piston 42g in the compression stroke, and the swirl is compressed and deformed in the axial direction, and a contraction of the swirl diameter occurs in the substantially spherical combustion chamber.
  • strong centrifugal separation occurs like the cyclone effect.
  • the fuel which is lighter than air and has a large density difference with air such as hydrogen or methane gathers to the center side in the cylinder axis symmetry by the strong centrifugal separation action, and the fuel is shown in FIG. 5 according to the density gradient of the gas.
  • the layers are separated into layers to form a high concentration combustible layer in the center and collected in the ignition part of the spark plug or injector provided near the intersection with the cylinder axis of the combustion chamber. Disturbance of swirl can be suppressed by providing a tubular valve recess 421 on the top surface of the piston 42g.
  • the centrifugal effect shortens the action time of centrifugal separation when the internal combustion engine is at high speed, but since the intake flow velocity increases and the swirl rotation becomes fast, the influence of the action by the rotational speed of the internal combustion engine is small.
  • FIG. 4 is a volume diagram at the end of exhaust (U1) and at the end of compression (U2) of the combustion chamber of the second embodiment (FIG. 2), and an explanatory view of the change of the radial volume occupancy by trial calculation of the compression stroke ( U3 (V)). Illustration and description of the swirling flow are omitted for the purpose of FIG.
  • Top left (U1) is a cross-sectional view of the combustion chamber where the radial dimension at the end of exhaust is divided into eight equal parts, and the left half flows toward the front and the right half flows toward the back by swirl. Is a substantially spherical combustion chamber, and the difference in volume occupancy between the peripheral portion and the central cylinder axis is gentle.
  • Top right (U2) is a cross-sectional view of the combustion chamber divided into eight equal parts in the radial direction at the end of compression, and the swirl is compressed, and the difference in volume occupancy between the central cylinder axis and the peripheral part increases .
  • the vertical axis is the occupancy rate
  • the horizontal axis is the cylinder diameter
  • the auxiliary line on the vertical axis is each diameter of concentric circles centered on the cylinder axis that divides the cylinder bottom area into four. .
  • the left figure (U3) in the figure below is the occupancy ratio converted from the cross-sectional view of the combustion chamber at the end of exhaust and at the end of compression, and is the change from the end of exhaust (dotted line) to the end of compression (solid line)
  • dotted line the change from the end of exhaust
  • solid line the change from the end of compression
  • the end of the compression (solid line) is obtained by converting the cross-sectional view (U2) of the combustion chamber at the end of the compression indicated by a two-dot chain line into an occupancy ratio.
  • the right figure (U3V) in the figure below is divided into four equal areas by concentric circles where the low area increases by 25% around the cylinder axis, so the estimated value of the increase or decrease of the occupancy rate of each area In the region), the compression stroke reduces the occupancy by half, while the occupancy of the (0 to 25%) region on the cylinder axis increases, so each region shifts toward the cylinder axis, and the swirl diameter of the swirl decreases. Therefore, the angular velocity of the swirl increases and the motor rotates at a high speed.
  • the bottom area of the cylinder is proportional to the square of the diameter, so even with a small amount of fuel according to the present invention There is an effect that the combustible layer can be formed with certainty.
  • the fuel having a density smaller than air moves upward, so that the fuel is further collected near the spark plug.
  • the shape of the combustion chamber changes from a substantially cylindrical shape to a substantially spherical dome shape in a compression stroke as viewed macroscopically, the laminar distribution of fuel in the radial direction of intake moves from the peripheral portion to the central portion, Since the swirling diameter of the swirl holds a kinetic energy to reduce the diameter, the angular velocity increases and the number of rotations increases, and a strong centrifugal separation occurs due to the increase of the centrifugal force as in the cyclone effect.
  • FIG. 5 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and high-speed flame propagation at the time of ignition in the second embodiment (FIG. 15).
  • FIG. 5 is a cross-sectional view of the TDC of Example 2 (FIG. 2) taken along the line XX, in which swirl generated by the tangential port 230 g causes centrifugal action on the premixed hydrogen, and hydrogen having a smaller density than air.
  • the fuel moves toward the center of the cylinder axis and collects near the cylinder axis provided with the spark plug 11g. As described in FIG.
  • the combustion action of the stratified distribution at the time of combustion stroke ignition of the internal combustion engine 1g of FIG. 5 forms a flame kernel in the high concentration layer (F1) by igniting the high concentration layer (F1) near the cylinder axis by the spark plug 11g. While diffusing high concentration fuel by thermal expansion due to flame propagation and combustion and approximately concentric swirl, the turbulent flame is uniformly propagated circumferentially from the inner side with high fuel concentration of each layer to the outer layer with low concentration.
  • the high concentration layer (F1) has a large separation speed because hydrogen of the fuel has a large density difference with air, and forms a high concentration layer of high concentration, and the combustion range (about 4 to 75% as shown in FIG. 6) ) Is large, you can do a reliable ignition.
  • the mixed diffusion of fuel and air (oxygen) by premixing is high concentration layer (F1), middle concentration layer (F2), low concentration layer ( F3)
  • Fuel mixing layer of ultra-low concentration layer (F4), flame propagation of combustion is low temperature combustion around the outer periphery of ultra-low concentration layer (F4) or flame extinction near wall surface, so knocking phenomenon is suppressed, Since the unburned fuel is reduced due to the extinction in the concentration layer F4, there is an effect of improving the combustion efficiency by suppressing the cooling loss of the internal combustion engine.
  • stratified charge combustion results in a lean burn engine with good combustion efficiency with less fuel than the prior art.
  • the reference numerals of the layers formed by centrifugation in the following examples are described using the same layer reference (F1 to F4) as that of the second embodiment (FIG. 5) for easy understanding.
  • gasoline supply it diffuses to the medium concentration layer (F2) or the low concentration layer (F3) depending on the supply timing, so the high-speed flame propagation of hydrogen promotes combustion, and the high-speed combustion of hydrogen raises the pressure in the combustion chamber
  • the SPCCI engine compression ignition combustion with spark control
  • the compression ignition of the fuel starts with a high concentration of hydrogen, the combustion is promoted, and the power increase effect is obtained by the increase of the rotational speed.
  • FIG. 6 is a characteristic diagram of fuel and air of the internal combustion engine etc. of the second embodiment (FIG. 2), wherein the fuel shows the combustion range and the density, and the air shows the composition ratio and the density.
  • the vertical axis of the figure is the density of gas (Kg / m 3), the horizontal axis is the combustion range of the fuel, and the composition ratio (Vol%) of the air.
  • the main composition ratio of air is about 21% oxygen and about 78% nitrogen, the density of oxygen is about 1.07 times that of air, the density of nitrogen is about 0.93 times that of air, and the density with air.
  • the differences between oxygen and nitrogen, which are small and diffused in the air, are difficult to separate in the short time centrifugation of the present invention.
  • the density of hydrogen which is a fuel
  • the density of methane is about 0.53 times that of air, which is not as easily centrifuged as hydrogen.
  • Fuel is separated in the innermost of the layers formed by density gradient by centrifugation. Since the density of propane is about 1.47 times that of air and the density of gasoline is about 2.71 times that of air, these fuels have a concentration layer outside of the innermost layer formed by density gradient in centrifugation.
  • the combustion range of the fuel on the horizontal axis of the fuel burns in a small fuel range except for hydrogen, but hydrogen ranges as large as 4.1 to 75 (Vol%) To burn.
  • the minimum ignition energy (mj) is less than hydrogen (0.02) than gasoline (0.24) and the maximum burn rate (cm / s) is less than hydrogen (346) than gasoline (42) Because it is large, hydrogen is suitable for combustion initiators because it has good ignition performance and high blast pressure. However, it has a problem that energy density is small because its calorific value is small as fuel.
  • FIG. 7 is a plan view of a three-cylinder internal combustion engine showing the arrangement of the reciprocating compressor, intake and exhaust valves and passages according to the third embodiment (corresponding to claim 1), and peripheral circuits such as scavenging air amplification means and positive displacement hydraulic pressure supply means.
  • FIG. FIG. 7 shows scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the exhaust gas amount in a two-stroke internal combustion engine 1j provided with an intake valve 46j and an exhaust valve 47j in a cylinder head.
  • the scavenging amplification means 5j is provided between the compressor 25j driven at the same time and the intake inflow path 22j and the intake and outflow path 23j, which are intake paths, and the scavenging amplification means 5j includes a check valve 55j and the check valve 55j.
  • the air flow amplifier 50j is provided downstream, the drive flow passage 58j of the air flow amplifier 50j is in communication with the discharge port of the compressor 25j, and the combustion chamber has a substantially spherical or substantially conical shape.
  • the intake valve 46 j and the exhaust valve 47 j are radially disposed in the cylinder, the spark plug 11 j is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port generates swirl in the cylinder
  • the fuel is supplied to the intake system and the combustion chamber, such as hydrogen and methane, which have a smaller density than air and fuel of a spark ignition type internal combustion engine, according to the operating condition of the internal combustion engine 1j.
  • the configuration of the scavenging gas supply means of each cylinder is the same as that of the first embodiment, and the reciprocating compressor 25j and the scavenging gas amplification means 5j are provided for each cylinder between the intake inflow passage 22j and the intake outflow passage 23j.
  • the intake and outflow passages 23j may be communicated by three cylinders to form an integral manifold, but the performance of scavenging and supercharging may be reduced.
  • the positive displacement hydraulic pressure supply means 8j can arrange the three hydraulic pressure passages 88 (j1 to j3) by sharing the cam 811 by adding the vanes 83j to the positive displacement hydraulic pressure supply means 8 of the first embodiment at equal intervals. Therefore, with a simple structure, the hydraulic means can be manufactured highly reliable, compact and inexpensive.
  • FIG. 8 is a cross-sectional view of the internal combustion engine 1j provided with cooling means for the intake valve and the exhaust valve of the JJ cross section of the third embodiment (FIG. 7).
  • FIG. 8 is a cross-sectional view of the intake valve 46j and the exhaust valve 47j which are gas exchange valves of the JJ cross section of FIG. 7, and the intake valve 46j is opened by the pressure of scavenging with a lift check valve.
  • the valve drive mechanism 47j is opened by the hydraulic pressure supplied from the hydraulic pressure passage 88j3 to the valve cylinder 471j, and the intake valve 46j is a lift check valve, and the exhaust valve 47j is a blow blade 476 integrated with the hydraulic piston.
  • a part of the intake air is sent from the communication pipe 464 to the contact space of the check valve of the intake valve and the valve cylinder 471j of the exhaust valve by the pump action of the blower blade to perform cooling.
  • the exhaust valve cylinder a portion of the intake air passing through the communication pipe 464 performs the cooling through the intake conduit 445.
  • FIG. 9 shows the construction of a two-stroke internal combustion engine provided with scavenging amplification means comprising a check valve and an air flow amplifier according to a fourth embodiment (corresponding to claim 1), and a reciprocating compressor whose phase is advanced from the crankshaft.
  • scavenging amplification means comprising a check valve and an air flow amplifier according to a fourth embodiment (corresponding to claim 1), and a reciprocating compressor whose phase is advanced from the crankshaft.
  • FIG. 9 shows, as scavenging air supply means, a reciprocating compressor 25d, which is a compressor driven by the internal combustion engine 1d, and a scavenging air amplification means 5 between the intake inflow passage 22d and the intake outflow passage 23d, which are intake passages;
  • the scavenging air amplification means 5 comprises a check valve 55 and an air flow amplifier 50 provided downstream of the check valve, and the drive flow passage 58 of the air flow amplifier 50 is a discharge port of the discharge valve 257d of the reciprocating compressor 25d. It is a two-stroke internal combustion engine 1d according to claim 1, which communicates with the engine.
  • the reciprocating compressor 25d whose phase is advanced by ⁇ d at a narrower angle than the output means 4d has a check valve in the suction valve 256d and the discharge valve 257d, and carries out a stroke by connecting the connecting rod 253d to the connecting rod 43d of the output means 4d.
  • the length of the cylinder 251d can be made smaller than the diameter of the cylinder 41d of the output means 4d, and the intake efficiency must be reduced by the scavenging air flow and the intake efficiency must be greater than the exhaust capacity.
  • Supercharging can be handled by the reciprocating compressor 25d with a small capacity by the flow amplification function of the air flow amplifier 50.
  • the suction valve 46 d of the output means 4 d is a check valve, and the drive wheel 401 provided on the crankshaft 44 d rotates the driven wheel 402 of the same effective diameter ( ⁇ D) via the transmission medium 403 and provided on the driven wheel 402
  • the exhaust valve 47 is opened and closed by the cam 408.
  • the heavy oil or light oil stored in the fuel tank 75 d of the fluid supply means 7 is pressurized by the supply pump 131 and supplied to the combustion chamber through the common rail 141 by the injector 12 d as appropriate.
  • the suction valve 256d of the reciprocating compressor 25d and the upstream of the scavenging air amplification means 5 communicate with the air cleaner 21d, and the exhaust valve 47d of the output means 4d communicates with the exhaust purification device 32d upstream of the silencer 33d via the exhaust passage 31d. .
  • the operation of the internal combustion engine 1d shown in FIG. 9 supplies compressed air generated by the reciprocating compressor 25 as a drive flow to the air flow amplifier 50 of the scavenging amplifier 5, and the air flow amplifier 50 amplifies the flow of intake air to increase
  • the intake air is supplied to the intake valve 46d of the output means 4d at a pressure higher than the atmospheric pressure to scavenge the internal combustion engine 1d. If the scavenging pressure on the downstream side of the air flow amplifier 50 becomes too high, the check valve 55 prevents the occurrence of a reverse flow amplification phenomenon in which the intake air flows back to the air cleaner 21d and the flow is amplified in the reverse direction by the drive flow. .
  • the check valve is activated and the drive flow directly flows into the scavenging air, so that the scavenging air has a high pressure and a supercharging action occurs. Since the flow rate of the intake air is amplified according to the flow rate amplification ratio of the air flow rate amplifier 50, the displacement of the reciprocating compressor 25d may be smaller than the displacement of the internal combustion engine 1d.
  • the diameter of the cylinder 251d of the reciprocating compressor 25d is smaller than the cylinder 41d of the output means 4d, and the stroke can be sufficiently scavenged with a small and inexpensive reciprocating compressor 25d having a short stroke, and the lubrication of the reciprocating compressor 25d is splashed by the output means 4d.
  • High reliability because lubrication can be shared.
  • Supercharging can be performed by complete gas exchange and pressurization of the scavenging air by the driving flow by supplying the scavenging air having a surplus by compensating the blow-by with the scavenging air amplification means 5 having a simple configuration.
  • the timing chart of the internal combustion engine 1 will be described with reference to FIG. 12, the in-cylinder pressure based on the timing chart and trial calculation will be described with reference to FIG. 13, and a PV diagram based on trial calculation at high speed rotation will be described with reference to FIG.
  • the internal combustion engine 1 of the fourth embodiment is a diesel engine, it may be a spark ignition internal combustion engine.
  • FIG. 10 shows a configuration example of the scavenging air amplification means of the fourth embodiment (FIG. 9), in which the flow transvectors (B) of the ejector (A) and the prior art (Japanese Patent Laid-Open No. 2016-125421) are arranged in ascending order of flow amplification ratio of the air flow amplifier.
  • the check valve may be a reed valve 551, a lift check valve 555 (C) or another check valve, which can be selected according to the specification of the internal combustion engine from the response, the pressure resistance, and the like.
  • the air flow amplifier is mainly selected by the flow amplification ratio, and when the operating condition of the internal combustion engine fluctuates, the pressure loss in the scavenging amplification means 5 increases in the high speed region etc. and the operation efficiency decreases.
  • the nozzle opening area variable type having a large operation area of the prior art Japanese Patent Application Laid-Open No. 2016-125421) shown in the above (c) is preferable.
  • FIG. 11 is an operation explanatory view of the scavenging amplification means 5 of the initial exhaust stroke (S1), scavenging stroke (S2), and compression stroke (S3) of the internal combustion engine 1d of the fourth embodiment (FIG. 9).
  • S1 initial exhaust stroke
  • S2 scavenging stroke
  • S3 compression stroke
  • the exhaust stroke (S1) the exhaust after the combustion is started to be exhausted by opening the exhaust valve 47d, and the reciprocating compressor 25 is at the initial stage of compression.
  • the supply flow rate amplification is started, but when the pressure of the exhaust is high, the intake valve 46d which is a check valve does not open.
  • the exhaust proceeds and the exhaust pressure decreases, and the driving flow pressure of the reciprocating compressor 25 rises, and the scavenging pressure of the intake outflow passage 23d rises by the flow amplification by the scavenging amplification means 5 and the intake
  • the flow rate amplification further proceeds to start scavenging of the cylinder 41d.
  • the compression stroke (S3) since the exhaust valve 47d is closed and scavenging air is supplied from the intake valve 46d, the cylinder 41d becomes higher than the atmospheric pressure and the pressure in the intake and outlet passage 23d increases.
  • the valve 55 is closed, and the compressed air of the reciprocating compressor 25 is directly supplied to the intake / outlet passage 23d and supplied from the intake valve 46d, so that a supercharging effect occurs.
  • FIG. 12 is a timing diagram of the internal combustion engine of the fourth embodiment (FIG. 9).
  • the combustion stroke (B) in FIG. 12 starts from the start of combustion at TDC and ends by opening the exhaust valve 47d whose phase is ⁇ E earlier than BDC, and the exhaust stroke (E) starts from opening the exhaust valve 47d.
  • the compression stroke ends with the closing of the exhaust valve 47d, and the compression stroke starts from the closing of the exhaust valve 47d and ends with the TDC.
  • the scavenging stroke starts from the middle of the exhaust stroke later by ⁇ S than BDC and ends in phase delayed by Cs from closing of the exhaust valve 47d, and the ⁇ s and Ds are reversely operated by the pressure difference between the exhaust and scavenging air.
  • the exhaust stroke does not have to be targeted with respect to the cylinder stroke by the operation of the exhaust valve 47d, so the combustion stroke can perform sufficient expansion work, and the scavenging stroke overlapping the exhaust stroke and the compression stroke is Exhaust gas is sufficiently discharged and scavenging air flows into the cylinder 41d reduced to substantially atmospheric pressure to perform efficient gas exchange, and by further supplying scavenging air after the exhaust valve 47d is closed, a supercharging effect is generated.
  • FIG. 13 shows in-cylinder pressures estimated by calculation and timing charts of respective parts of an internal combustion engine provided with the reciprocating compressor and the scavenging air amplification means of the fourth embodiment (FIG. 9).
  • the horizontal axis of FIG. 13 is the crank angle displacement amount of the two-stroke internal combustion engine (360 °), and each item of the vertical axis is the time chart (band graph) of the internal combustion engine 1d of FIG.
  • Next is a timing chart of fuel supply (different from the fourth embodiment (including spark ignition type internal combustion engine) and (including scavenging air mixing)), next is a timing chart of each element of the output means 4d and the reciprocating compressor 25d.
  • the hatched portion of the variation of the piston 252d of the machine 25d indicates the suction amount of the reciprocating compressor 252d.
  • the lowermost portion is a trial calculation value of the in-cylinder pressure fluctuation of the cylinder 41d that generates the output of the internal combustion engine 1d as a result of the above.
  • the fuel supply can prevent the outflow of unburned fuel due to scavenging air by fueling the internal combustion engines at the timing of hatching.
  • the fluctuation of the pressure in the lowermost cylinder 41d is the combustion (combined cycle) at high speed and high load of the internal combustion engine 1d which is a diesel engine, and even if it is a spark ignition type internal combustion engine, except for FIG. The same is true for the timing chart, although the scavenging timing slightly fluctuates due to the supercharging pressure and the like, but there is no big difference.
  • FIG. 14 is a PV diagram based on a trial calculation at high speed rotation of an internal combustion engine provided with the scavenging air amplification means of the fourth embodiment (FIG. 9).
  • the vertical axis of the PV diagram of FIG. 14 is the in-cylinder pressure P, which is the same absolute pressure (abs) as the in-cylinder pressure of the cylinder 41 d of FIG. 13.
  • the horizontal axis is between the BDC and TDC of the internal combustion engine 1d.
  • Stroke volume Vst due to piston movement of the FIG. 14 is a PV diagram of a combined cycle based on a trial calculation of the internal combustion engine 1 d which is a diesel engine, but the combustion cycle is different from that of the spark ignition type internal combustion engine.
  • the thick lines (points E4 to E3) in the figure are PV diagrams of the internal combustion engine 1d during high speed and high load operation, and the exhaust stroke is from point E3 at which the exhaust valve 47d opens to piston E 42d at point E4.
  • the compression stroke is from the point ES6 through the point S7 to the point C1 of the TDC while the piston 42d is moving to the TDC side, and the combustion stroke is At the point C1, the TDC is turned back to pass through the point B2 to the point E3.
  • the scavenging stroke is from the point ES5 in the exhaust stroke to the point S7 of the compression stroke through the point ES6 at which the exhaust valve 47d closes.
  • FIG. 15 is a cross-sectional view of an internal combustion engine provided with scavenging amplification means provided with a lift check valve according to the fifth embodiment (corresponding to claim 1) and a transvector according to the prior art (Japanese Patent Laid-Open No. 2016-125421) and a reciprocating compressor. It is.
  • FIG. 15 shows, as scavenging air supply means, a reciprocating compressor 25u which is a compressor driven by the internal combustion engine 1u, and a scavenging air amplification means 5u provided between an intake air inflow passage 22u and an intake air outflow passage 23u which are intake air passages.
  • the scavenging air amplification means 5u comprises a lift check valve 555u and a transformer vector 53u of the prior art (Japanese Patent Laid-Open No. 2016-125421) which is an air flow amplifier provided downstream of the lift check valve 555u.
  • the two-stroke internal combustion engine 1u according to claim 1, wherein the drive flow passage 58u of the vector 53u is in communication with a discharge valve 257u which is a discharge port of the reciprocating compressor 25u.
  • the drive flow passage 58u is provided with cooling fins to air-cool the adiabaticly compressed drive flow, but may be shared with liquid cooling and cooling such as a reciprocating compressor 25u.
  • the intake valve 46u is a check valve, and the exhaust valve 47u supplies hydraulic pressure generated by a positive displacement hydraulic pressure supply means 8u (not shown) from the hydraulic passage 88u to the valve cylinder 471u by the pressure difference between the intake flow passage 23u and the combustion chamber. Open and close.
  • the connecting rod 253u of the reciprocating compressor 25u is piggyback connected to the connecting rod 43u of the output means to advance the phase and shorten the stroke, and the cylinder 251u of the reciprocating compressor 25u is smaller than the cylinder 41u of the output means and splashed It can share lubrication.
  • valve drive of the exhaust valve 47u is the same as that of the first embodiment, and the operation of the reciprocating compressor 25u and the scavenging amplification means 5u is the same as that of the fourth embodiment and the explanation is omitted. This will be described in 16.
  • FIG. 16 is a cross-sectional view of the scavenging amplification means 5u configured of a variable nozzle type transvector and a lift check valve according to the prior art (Japanese Patent Laid-Open No. 2016-125421) of the fifth embodiment (FIG. 15).
  • the transformer vector 53u of FIG. 16 is provided with a nozzle adjustment mechanism consisting of a spring 535 for urging the piston 534 provided with the nozzle surface on the upstream side of the nozzle 531 in the nozzle closing direction on the inner peripheral surface of the housing 533
  • a lift check valve 555 u consisting of a disk 557 and a spring 556 is provided between the piston 534 and the flange 536 of the nozzle adjustment mechanism.
  • the lift check valve 555 u urges the disc 557 and the disc 557 to a bearing surface of the cylinder portion by a spring 556 against a cylinder portion provided on the flange 536, and the disc 557 is closed when seated on the seat surface.
  • a plurality of communication ports, a contact for restricting the stroke at the substantially ring-shaped outer peripheral end, and a guide convex portion at the central portion for smoothing the intake flow and reducing the passage resistance are provided.
  • the action of the scavenging air amplification means 5u is that the disk 557 is urged to the bearing surface of the flange 536 by the urging force of the spring 556 and the backflow air intake at the time of the backflow of the air intake generated by the pressure increase of the intake air downstream by the lift check valve 555u.
  • the lift check valve 555u is opened and the intake air supplied from the intake inflow passage 22u To the air outflow passage 23 u.
  • the scavenging air amplification means 5u amplifies the scavenging air flow rate by controlling the flow rate of the driving flow according to the driving flow pressure and scavenging flow rate conditions by the nozzle adjustment mechanism of the transformer vector 53u, and the reverse flow rate amplification phenomenon by the lift check valve 555u. As shown in (S3) of FIG. 11 of the fourth embodiment, direct supercharging is performed by the drive flow.
  • FIG. 17 is a plan view (1), a sectional view taken along the line KK (2), L of a structural explanatory view of a two-cylinder internal combustion engine 1k provided with the scavenging air amplification means of the sixth embodiment (corresponding to claim 1) and a reciprocating compressor.
  • FIG. 6 is a cross-sectional view (3). As shown in the plan view of the two-cylinder internal combustion engine 1k in the upper diagram (1) in FIG.
  • a reciprocating compressor 25 (k, L) is provided in each cylinder whose combustion chamber is substantially spherical,
  • the drive flow from the reciprocating compressor 25 (k, L) is supplied from the drive flow passage 58k, and the scavenging air is amplified from the scavenging amplification means 5k to the tangential port of each cylinder and supplied to the scavenging of the internal combustion engine 1k. And supercharge.
  • the crank angle of each cylinder differs by 180 °, so the scavenging operation timing does not interfere as shown in FIG.
  • each reciprocating compressor 25 (k, L) and the scavenging amplification means 5 k is the same as that of the fourth embodiment. Since the scavenging air amplification means 5k can be shared, the internal combustion engine 1k can be manufactured inexpensively with a simple configuration, has a high combustion efficiency, and has a high output.
  • FIG. 18 is a timing chart of each cylinder of the internal combustion engine of the sixth embodiment (FIG. 17) and in-cylinder pressure based on trial calculation.
  • FIG. 18 describes the data of the two-cylinder (K-K) (L-L) in each item, but the chart creation method is the “timing of the internal combustion engine 1 d of the fourth embodiment (FIG. 13) Since it is the same as the “in-cylinder pressure based on the chart and trial calculation”, the description of the method of creating the chart is omitted.
  • the thick solid line in the figure is the first cylinder (K-K), the thick broken line (L-L) is the second cylinder, and the intake and outflow passages 23k are open by the combustion chambers of the respective cylinders and the intake valves (k, L).
  • the scavenging stroke of each cylinder does not interfere because the crank angle is 180 ° out of phase or not, as can be seen in the operation strokes of the figure. , L) and ejection timing do not interfere.
  • FIG. 19 is a plan view (1) of a configuration explanatory view of a three-cylinder internal combustion engine provided with the scavenging air amplification means of the seventh embodiment (corresponding to claim 1) and a reciprocating compressor, and an MM cross section of the plan view (1). It is a figure (2).
  • the internal combustion engine 1m is provided with reciprocating compressors 25 (m1 to m3) parallel to each cylinder of the in-line three-cylinder output means whose phase difference between the cylinders is 120.degree.
  • the intake flow passage 23m is in communication with all the intake valves (tangential ports) of the in-line three cylinders.
  • the discharge of the reciprocating compressor 25 partially buffers but there is no effect on many, and the cylinders of the internal combustion engine 1m can be compacted.
  • a large internal combustion engine can be used, and the other advantages such as the flammability are the same, so the description will be omitted.
  • FIG. 20 is an explanatory view of a configuration concept of a two-stroke internal combustion engine 1s in which a control valve 56 is provided in the reciprocating compressor 25s of the eighth embodiment (corresponding to claim 1) and the scavenging amplification means 5s.
  • a control valve 56 is provided in the reciprocating compressor 25s of the eighth embodiment (corresponding to claim 1) and the scavenging amplification means 5s.
  • a reciprocating compressor 25s driven by the internal combustion engine 1s and a scavenging air amplification means 5s are provided in the intake passage 58s, and the scavenging air amplification means 5s includes the check valve 55s and the check valve
  • a control valve 56 is provided upstream of the air flow rate amplifier 50s, and the control valve 56 is controlled according to the operating condition of the internal combustion engine 1s to adjust the scavenging pressure and flow rate.
  • an ignition plug 11s is provided at the center of the combustion chamber of the internal combustion engine 1d of the fourth embodiment (FIG. 1) and the main fuel is supplied to the combustion chamber 12s.
  • This is a spark ignition internal combustion engine 1 s provided with an injection 12 s 2 for supplying a gaseous fuel whose density is smaller than that of air.
  • the exhaust gas recirculation passage 36 communicating with the exhaust gas passage 31s and the drive flow passage 58s is provided, and backflow to the control valve 38 and the exhaust gas passage 31s is prevented in the exhaust gas recirculation passage 36. It is possible to provide the check valve 37 to enable the EGR that can use the exhaust pressure as the drive flow pressure.
  • scavenging air is amplified at the air flow rate amplifier 50s of the scavenging air amplification means 5s by a drive flow supplied from the reciprocating compressor 25s, and scavenging is supplied by compensating for blowby.
  • supercharging is performed by complete gas exchange and pressurization of scavenging air by drive flow.
  • the control valve 56 By adjusting the control valve 56 to control the intake air temperature at the TDC of the combustion chamber etc.
  • the scavenging amplification means 5s performs complete scavenging and supercharging with a large amount of scavenging by normal flow rate amplification, and performs work for the outside in a combustion cycle like a two-dot chain line in the PV diagram shown in FIG.
  • the supercharging effect due to the pressure increase in the scavenging stroke is suppressed, as shown by the thick line of the PV diagram.
  • the control valve 56 may be a butterfly valve having a simple motorized structure, or may be a poppet type capable of linear opening control.
  • FIG. 21 is a PV diagram based on a trial calculation at the time of suppression of the flow rate amplification of scavenging by the control valve 56 provided in the scavenging amplification means of the eighth embodiment (FIG. 20).
  • the drawing method of FIG. 21 is the same as that of the fourth embodiment (FIG. 14), so the description will be omitted. Scavenging and supercharging can be adjusted by the control valve 56 of the exhaust amplification means 5s, and the thick lines (point E4b to point E3b) in FIG. 21 indicate supercharging action by control of the control valve 56 during low load operation of the internal combustion engine 1s.
  • the supercharging by the scavenging air amplification means 5 increases the output of the internal combustion engine 1s, so downsizing of the internal combustion engine can be performed, and the control valve 56 provided in the scavenging air amplification means 5 corresponds to the operating condition of the internal combustion engine 1s. Since the temperature of the premixed gas at TDC can be adjusted by scavenging and supercharging, the operation can be switched to an HCCI engine or an SPCCI engine.
  • the internal combustion engine 1d according to the fourth embodiment (FIG. 14) is a diesel engine
  • the internal combustion engine 1s according to the eighth embodiment shown in FIG. 21 is a spark ignition internal combustion engine.
  • the theoretical cycles (Sabati cycle and Otto cycle) are different. However, both embodiments do not limit the type of internal combustion engine, and may be a diesel engine or a spark ignition internal combustion engine.
  • FIG. 22 a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber of the ninth embodiment (corresponding to claim 2), and exhaust valves are opened and closed by cams provided on each of two camshafts. It is explanatory drawing of the construction concept of the 2 cylinder 2 cycle internal combustion engine which provided the scavenging air amplification means which makes a valve a non-return valve.
  • a plurality of intake valves 46n and exhaust valves 47n are alternately arranged radially in the combustion chamber 410n, and two parallel camshafts 407 (-1, -1, interlocked at the same rotational speed as the rotational speed of the crankshaft 44n).
  • the intake valve 46 n is a lift check valve, and is opened by the pressure difference between the scavenging air and the combustion chamber.
  • the internal combustion engine 1n drives a driven vehicle 402 provided on the crankshaft 44n via a transmission medium 403n to drive a driven vehicle 402 having the same effective diameter ( ⁇ Dn) as the drive vehicle 401 provided on the camshaft 407-1. Synchronize the rotation of -1 to the crankshaft 44n.
  • a cam 408 (-1, -3) is provided by a driven gear 406 having the same pitch circle diameter as the drive gear 405 provided on the cam shaft 407-2 meshing with the drive gear 405 provided on the cam shaft 407-1.
  • the cam shaft 407-1 provided with the cam shaft 407-1 and the cam 408 (-2, 4) rotates in the opposite direction at the same rotation speed.
  • the cams 408-1 and 408-2 have cam shapes that are symmetrical with respect to the cylinder axis, and the exhaust valves 47n operated by the respective cams open in synchronization with the crankshaft 44n.
  • a plurality of intake valves 46n and exhaust valves 47n are alternately arranged radially in the combustion chamber 410n, and the intake valves 46n and the exhaust valves 47n are disposed on the same line, and the intake valves 46n and the exhaust valves 47n as shown in the figure below.
  • the operation of the fuel supplied from the tangential port 230n and the injector 12n of the internal combustion engine 1n is the same as that of the first embodiment, and the operation of the reciprocating compressor 25n and the scavenging amplification means 5n is the same as that of the eighth embodiment.
  • FIG. 23 is a plan view and peripheral circuits of a two-cylinder two-stroke internal combustion engine in which the exhaust valve is cam driven and the intake valve is hydraulically driven via hydraulic means by two camshafts of the tenth embodiment (corresponding to claim 2)
  • FIG. 23 a plurality of intake valves 46p and exhaust valves 47p are alternately arranged radially in the combustion chamber, the drive wheel 401p is for the crankshaft 44p, and the driven wheel 402p with the same effective diameter as the drive wheel 401p is for the camshaft 407p.
  • a gear 405p is provided, and the cam shaft 407p2 is provided with a driven gear 406p having the same pitch circle diameter that meshes with the drive gear 405p, and two parallel cam shafts interlocked at the same rotation speed as the rotation speed of the crankshaft 44p ( 407p, 407p2), and the exhaust valve 47p is opened and closed by a cam 408p provided on each of the two cam shafts (407p, 407p2), and the intake valve 46p is operated by hydraulic means interlocking with a cam other than the cam
  • the two-stroke internal combustion engine 1p according to claim 1, which is opened and closed by a valve drive unit 80 (p1, p2, p3).
  • the driving wheel 401p drives the driven wheel 402p via the transmission medium 403p, and the valve driving unit 80 (p1, p2, p3) is a hydraulic pressure generated by the plunger 84 (p1, p2, p3) interlocking with the cam. Open the intake valve 46p.
  • the actions of the fuel supplied from the tangential port 230n of the internal combustion engine 1p and the injector 1p are the same as those of the first embodiment, and the actions of the reciprocating compressor 25p and the scavenging amplifier 5p are the same as those of the eighth embodiment.
  • FIG. 24 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the tenth embodiment (FIG. 23).
  • FIG. 24 is an explanatory view of a state where a piston 47g provided with a spherical cavity 420p of radius SRp2 at the center of the piston top surface in the above-mentioned Embodiment 10 (FIG. 23) is ignited by the spark plug 11p at TDC.
  • the valve is opened by the hydraulic pressure supplied from the valve drive unit 80p2, and the exhaust valve 47p is opened by a cam 408p provided on the cam shaft 407p.
  • the hydrogen ignited by the spark plug 11p propagates at a high speed from the high concentration layer F1p and burns, and burns together with a part of the gasoline diffused to the high concentration layer side F1p.
  • the SPCCI engine compressed compression auto-ignition
  • the HCCI engine can also be obtained by controlling the temperature of the premixed gas at TDC above the ignition point (300 ° C.) of gasoline by controlling the control valve 56p. As shown in FIG.
  • the annular thin layer with low fuel concentration is mainly in contact with the spherical cavity 420p on the top face of the piston by the above action, and the combustion of the high concentration combustible layer (F1p) is hindered by the piston 42p. Combustion is improved and most of the area on the top surface of the piston 42p is in contact with the low concentration layer (F3p, F4p), so the heat loss is suppressed and the output is improved. is there.
  • the action of the centrifugal separation action and the like by the swirl is the same as that of the second embodiment, so the description will be omitted.
  • FIG. 25 is an explanatory view of a configuration concept of an internal combustion engine using hydrogen as a fuel and provided with the scavenging air amplification means and the oxyhydrogen generator according to Embodiment 11 (corresponding to claim 3).
  • FIG. 25 shows the internal combustion engine 1h provided with an oxyhydrogen generator 9 operated by electrical means, and supplying the oxyhydrogen generated by the oxyhydrogen generator 9 as a fuel having a density smaller than that of the air. 2 cycle internal combustion engine 1h.
  • the hydrogen stored under pressure in the fuel tank 75h of the fluid supply means 7h is depressurized by the pressure reducing valve 64h of the fluid control means 6h, and the oxyhydrogen generation device 90 of the oxyhydrogen generation means 9 is operated with the secondary battery 96 to make the electrolyte
  • the amount of oxygen and hydrogen generated by electrolyzing the electrolytic solution in the tank 94 is adjusted by the control valve 63h of the fluid control means 6h, and the hydrogen and the acid hydrogen are supplied to the high pressure fuel pump 13h to pressurize them.
  • the mixing ratio of hydrogen and oxyhydrogen is adjusted by the fuel sensor 62h to determine the fuel injection amount.
  • the oxyhydrogen generator 90 can also be operated by the electric power generated by the generator driven by the internal combustion engine 1h. Since oxyhydrogen can be generated by electrolysis, it is possible to reduce the self-supply of hydrogen or the replenishment amount of hydrogen in the fuel tank 75h, and the hybrid vehicle as shown in Example 13 using the secondary battery 96 of the oxyhydrogen generator 90. Energy regeneration is possible.
  • the structure of the valve drive mechanism of the exhaust valve 47h, the configuration of the reciprocating compressor 25h, and the like, and the operation are the same as those of the ninth embodiment, and thus the description thereof will be omitted.
  • FIG. 26 is an explanatory view of a configuration concept of an oxyhydrogen generator for adding an ultrasonic wave to an electrolytic solution according to the prior art (Patent Document 9) of Example 12 (corresponding to claim 3). And the like.
  • FIG. 26 shows a cathode 911 and an anode 912 arranged in layers in the electrolytic cell 914, a DC power supply 913 for applying a DC voltage between the cathode 911 and the anode 912, and an electrolyte for controlling the supply of the electrolyte 915.
  • An electrolysis means 91 comprising a pump 917 as control means and a control valve 918 as gas collection means for collecting gas generated by electrolysis, an ultrasonic oscillator 921 and the ultrasonic oscillation
  • an ultrasonic oscillation means 92 comprising: a high frequency generator 922 for ultrasonically vibrating the element 921 by electrical means; and an ultrasonic wave propagating through the ultrasonic vibrator and the electrolytic solution
  • the distance between the acoustic wave and the reflection surface is an integral multiple of a quarter of the ultrasonic wavelength
  • the electrolysis means 91 is an ultrasonic wave that oscillates the cathode 911 and the anode 912 from the ultrasonic wave generator 921
  • the cathode 911 and the anode 912 are slat-like or grid-like so that ultrasonic oscillation can be propagated through the cathode 911 and the anode 912 in a plane perpendicular to the propagation direction of the oscillation.
  • FIG. 27 is an illustration of a concept of a spark-ignition internal combustion engine of a hybrid vehicle provided with a scavenging air amplification means, an oxyhydrogen generator and a regeneration means according to a thirteenth embodiment (corresponding to claim 3).
  • the intake port of the internal combustion engine 1s of the eighth embodiment is a tangential port 230t
  • the combustion chamber is substantially spherical with a radius SRt
  • the injector 12t is not a drive flow passage.
  • the acid hydrogen supplied to the injector 12t controls the amount of generated acid hydrogen by controlling the acid hydrogen generator 90t of the acid hydrogen generating means 9t, and the pressure of the generated acid hydrogen is adjusted by the control valve 63 of the fluid control means 6.
  • the excess oxyhydrogen is stored / released to the accumulator 67 at a control valve 63-2 in a timely manner, the oxyhydrogen is stably supplied, the regenerative energy is converted to the oxyhydrogen, and stored. Reduce the load.
  • the oxyhydrogen is timely supplied from the injector 12t to the intake and outflow passage 23t, and premixed with the intake air flowing into the combustion chamber after the exhaust valve 47t is closed, thereby preventing the blowout of fuel.
  • Gasoline which is the main fuel of the spark ignition type internal combustion engine 1t to be supplied to the injector 12t2 is stored in the fuel tank 75t of the fluid supply means 7t, pressurized by the high pressure fuel pump 13t and supplied, and the injector 12t2 discharges the exhaust valve 47t. Immediately after the valve is closed, it is supplied to the combustion chamber to prevent the blowout of fuel. In the fuel supplied to the combustion chamber, hydrogen is collected in the vicinity of the spark plug 11t by the centrifugal separation action of the swirl as in the first embodiment, and hydrogen having a small minimum ignition energy is ignited by the spark plug 11t.
  • the uniform flame propagation in the circumferential direction from the center of the cylinder axis promotes the combustion of the main fuel to increase the output of the internal combustion engine 1t. Since the outermost layer of the cylinder-axisymmetric layered fuel layer is a very low concentration layer, it is possible to prevent knocking that the remaining unburned mixture burns rapidly before reaching the flame propagation, and the outermost layer is on the combustion chamber wall Although the area in contact is large but because it is a very low concentration layer, cooling loss can be suppressed, and oxygen of oxyhydrogen has the effects of improving thermal efficiency and improving exhaust properties by oxygen enrichment of intake air.
  • the oxy-hydrogen generating unit 9t is connected in parallel to a secondary battery 96t for supplying DC power and an inverter 97, and sends and receives electric energy to the motor / generator 98 by the inverter 97 to assist the output of the internal combustion engine 1t. Or perform energy regeneration.
  • the control system of the internal combustion engine 1t is an explanatory view of a configuration concept of a control system (FIG. 27) described later, and a control flowchart of the internal combustion engine 1t such as operation switching of the hybrid vehicle to HCCI engine or SPCCI engine will be described later This will be described with reference to FIG.
  • FIG. 28 is an explanatory view of a configuration concept of a control system which can be operated as an HCCI engine or an SPCCI engine in the internal combustion engine of the hybrid vehicle of the thirteenth embodiment (FIG. 27).
  • the ECU 19 which is an electronic control unit for the internal combustion engine 1t is composed of a CPU (central processing unit), a storage element consisting of a RAM and a ROM, an input port, an output port, and a DC power source.
  • the relay device (controller, amplifier, converter, etc.) is not shown.
  • the internal combustion engine 1t of the hybrid vehicle includes input information such as a crank angle sensor 45, a knock sensor 48, a water temperature sensor 49, etc., an accelerator opening sensor 17 which is a control assisting device of the hybrid vehicle, a brake opening sensor 18,
  • the ECU 19 analyzes, judges, and predicts the operating condition of the internal combustion engine 1t based on input information from the vehicle speed sensor etc., and adapts the spark plug 11t and the injector (12t, 12t2) of the output means 4t to the operating condition of the internal combustion engine
  • the control valve 56t of the scavenging amplification means 5t adjusts the scavenging amount and the supercharging pressure
  • the control valve (63, 63-2) of the fluid supply means 6 adjusts the supply pressure of oxyhydrogen to the exhaust gas recirculation passage.
  • the exhaust gas recirculation amount is adjusted and controlled by the control valve 38t provided.
  • the control system of the internal combustion engine 1t causes the ECU 19 to control the internal combustion engine 1t according to the following control flowchart (FIG. 28) in accordance with the driving condition of the hybrid vehicle.
  • FIG. 29 is a control system (FIG. 28) of an internal combustion engine of the hybrid vehicle incorporating the internal combustion engine 1t (FIG. 27) of the thirteenth embodiment, in compression ignition combustion by HCCI engine or spark control of homogeneous charge compression auto ignition. It is a control flow chart (except motor control, such as drive assist) which operates as a SPCCI engine.
  • FIG. 29 is controlled by the ECU 19 that processes input / output information of the control system (FIG. 28).
  • the determination of acceleration or deceleration is made mainly by the driver's intention or the internal combustion engine 1t by input information from the accelerator opening sensor 17, brake opening sensor 18 and vehicle speed sensor (not shown) by accelerator or brake pedal operation.
  • the ECU 19 determines whether a combustion operation for acceleration is required by the accelerator opening sensor 17, the brake opening sensor 18, the vehicle speed sensor, etc. (step S10).
  • the HCCI engine or the SPCCI engine is predicted from the supercharge pressure of the internal combustion engine 1t, the water temperature of the internal combustion engine, etc. It is determined whether the operation is possible (step S11).
  • step S12 when it is determined that the combustion operation is not necessary, it is determined whether the brake opening degree sensor 17 is ON (step S12).
  • step S13 if it is determined that the brake opening sensor 17 is ON, it is determined whether energy regeneration is possible (step S13). Specifically, whether or not energy regeneration is possible is determined based on the kinetic energy that can be regenerated by the vehicle speed sensor or the like and the estimated amount of deceleration.
  • step S14 an inertial operation (free run) subroutine not actively accelerating or decelerating is executed (step S14), and this processing routine is temporarily ended at RETURN. .
  • step S13 If it is determined that energy regeneration is possible in the determination of whether the energy regeneration is possible (step S13), the energy regeneration subroutine (S15) is executed, and the electric power generated by the motor / generator 98 by the inverter 97 Is supplied to the oxyhydrogen generator 90t and the secondary battery 96t to perform energy regeneration, and the torque by the back electromotive force of the motor / generator 98 is used for braking.
  • an engine brake subroutine step S16 is executed. Specifically, the fuel supply is stopped, and the control valve 56 of the scavenging amplification means 5t is controlled so that compression work, pumping loss and the like occur.
  • step S11 If it is determined in step S11 that HCCI operation or SPCCI operation is not possible based on the pressure and temperature prediction of the intake air of the combustion chamber at TDC, the control valve 56t of the scavenging gas amplification means 5t, the exhaust gas recirculation passage Execute the engine operation adjustment subroutine (step 17) so that HCCI operation or SPCCI operation is possible by controlling the control valve 38 etc. of 36 and operating the combustion chamber etc. of the internal combustion engine 1t, and determine whether the main fuel is necessary. (Step 21). If it is determined in the step 11 that the HCCI operation or the SPCCI operation of the internal combustion engine 1t is possible, it is determined whether the hydrogen fuel can be supplied (step 18).
  • the HCCI or SPCCI operation subroutine (step 19) by the main fuel of the internal combustion engine 1t is executed, and the processing routine is temporarily ended at RETURN.
  • the control valve 56t of the scavenging gas amplification means 5t, the control valve 38 of the exhaust gas recirculation passage 36, and the like are controlled to make the intake temperature at TDC of the internal combustion engine 1t equal to or higher than the ignition point of the main fuel, or the intake temperature And adjust the operation so as to ignite with the spark plug 11t near the ignition point of the main fuel.
  • step 18 the HCCI or SPCCI operation subroutine (step 20) is executed using the hydrogen fuel of the internal combustion engine 1t or two fuels of the main fuel and hydrogen fuel. After that, this processing routine is temporarily ended at RETURN. If it is judged at step 21 that the main fuel is necessary, then it is judged whether hydrogen fuel can be supplied if it is judged that the main fuel is necessary (step 22). On the other hand, if it is determined that hydrogen fuel can be supplied, a spark ignition type internal combustion engine operation subroutine (step 24) is executed by the two fuels of the main fuel of the internal combustion engine 1t and hydrogen fuel, and then the processing routine is executed at RETURN. Once.
  • a spark ignition type internal combustion engine operation subroutine (step 25) using the hydrogen fuel of the internal combustion engine 1t is executed, and this processing routine is temporarily ended at RETURN.
  • fuel supply, supercharging, EGR, energy regeneration and the like of the internal combustion engine 1t are executed according to each subroutine according to the operating condition of the spark ignition internal combustion engine 1t of the hybrid vehicle.
  • the control system (FIG. 24) of the internal combustion engine 1t of the hybrid vehicle is controlled by the input / output information of the ECU 19, and this control flowchart is repeatedly executed during operation of the internal combustion engine 1t.
  • FIG. 30 is an explanatory view of a construction concept of a four-stroke internal combustion engine for opening and closing each intake valve, the upper view is a plan view of the internal combustion engine 1a, and the lower view is a cross section of the internal combustion engine 1a FIG. In FIG.
  • the combustion chamber has a substantially spherical shape with a radius SRa, a plurality of intake valves 46a and exhaust valves 47a are arranged alternately in the combustion chamber, and the injector 12d is of the combustion chamber.
  • the intake port is provided near the intersection with the cylinder axis, and the intake port is a tangential port 230a that generates swirl in the cylinder, and fuel such as hydrogen having a density smaller than that of air is supplied to the combustion chamber to operate the internal combustion engine 1d
  • fuel such as hydrogen having a density smaller than that of air
  • the supply of the fuel is controlled, and further, two parallel camshafts 407 (a1, a2) interlocking at a rotational speed of 1/2 of the rotational speed of the crankshaft 44a are provided, and the exhaust valve 47a is A pump which is a hydraulic means that opens and closes by means of cams 408 (a1, a2) provided on the two respective camshafts 407 (a1, a2) and interlocks the intake valve with the cams 408 (a1, a2) Nja 84a (a1, a2), the valve cylinder 471 (a1, a2), is a 4-cycle internal combustion engine 1a for opening and closing the valve piston 472 (a1, a2).
  • the internal combustion engine 1a drives a driven vehicle 402a having an effective diameter (.phi.2Da) twice the effective diameter .phi.Da of the drive vehicle 401a provided on the camshaft 407a1 via the transmission medium 403a by the drive vehicle 401a provided on the crankshaft 44a.
  • the cam shaft 407a1 is rotated at a half rotation speed of the crank shaft 44a.
  • the cam shaft 407a2 provided with the cams 408 (a2, a4) rotates in the opposite direction at the same rotational speed.
  • the cams 408a1 and the cams 408a2 have cam shapes symmetrical with respect to the cylinder axis, and the exhaust valves 47a operated by the respective cams are opened in synchronization with the crankshaft 44a.
  • the intake valve 46a is a hydraulic pressure generated by a plunger 84 (a1, a2) that operates with a phase lag behind the exhaust valve 47a provided to the cam 408 (a1, a2).
  • the valve piston 472 (a1, a2) is operated to open and close the valve.
  • a plurality of intake valves 46a and exhaust valves 47a are alternately arranged radially in the combustion chamber, and the intake valves 46a and the exhaust valves 47a are arranged on the same line, and as shown in the figure below, the intake valves 46a and the exhaust valves 47a are By setting the narrow angle to ⁇ a ( ⁇ a> 90 ⁇ ), the arrangement interference of multi-cylinder valves can be reduced, the distance between cylinders can be shortened, and a compact, lightweight, rigid cylinder block can be obtained.
  • the centrifugal separation action of fuel having a density smaller than air, such as hydrogen supplied from the tangential port 230a of the internal combustion engine 1a and the injector 12a has the same basic principle as the first to third embodiments of the two-stroke internal combustion engine.
  • the intake valve 46a needs an exhaust residual pressure that is substantially atmospheric pressure by the exhaust valve 47a and a valve thrust that resists the spring 473a, but may have a valve thrust that is smaller than the valve thrust of the exhaust valve.
  • FIG. 31 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and combustion of an HCCI engine at the time of fuel injection in the fourteenth embodiment (FIG. 30).
  • FIG. 31 is a distribution state of a fuel concentration layer at TDC of hydrogen, which is a premixed fuel supplied from the injection 12a, by the centrifugal separation action of the fuel having a density smaller than that of the air in the fourteenth embodiment (FIG. 30).
  • the diesel fuel which is the main fuel injected from the injection 12 a near the intersection of the combustion chamber and the cylinder axis, is a hydrogen premixed gas that is adiabatically compressed at TDC and at a temperature above the ignition point (250 ° C.) of the diesel.
  • the cylinder shaft rotating at high speed by swirling
  • the high temperature and high pressure flame propagation from the center expands uniformly in the circumferential direction, and hydrogen in the middle concentration layer (F2a) also starts combustion by the combustion of the light oil and the flame propagation, so the pressure and temperature of the hydrogen in the combustion chamber increase.
  • Combustion in the entire combustion chamber is promoted by igniting hydrogen in a region where hydrogen combustion is diffused, such as low concentration layer (F3a), exceeding the ignition point (585 ° C), and generation of particulates, deposits, etc. Suppress.
  • the outer circumferential layer has less generation of unburned gas due to the uniform combustion in the circumferential direction, suppresses knocking in the ultra-low concentration layer (F4a) of premixed hydrogen as the outer circumferential layer, and has a large contact area with the wall surface of the combustion chamber. Since the amount of heat generation of (F4a) is small and the cooling loss is small, there is an effect of improving the thermal efficiency of the internal combustion engine 1a.
  • FIG. 1 is a plan view and a peripheral circuit diagram of a four-stroke internal combustion engine.
  • the combustion chamber has a substantially spherical shape with a radius SRb, and a plurality of intake valves 46 (b1 to b4) and exhaust valves 47 (b1 to b4) are alternately arranged radially to the combustion chamber.
  • the four-stroke internal combustion engine 1b is opened and closed by a valve drive unit 80 (b1, b3) which is hydraulic means interlocking with a cam different from the cam 408.
  • a rotational drive that decelerates to a half of the drive wheel 401b provided on the crankshaft 44b and the driven vehicle 402b provided on the camshaft 407b1 via the transmission medium 403b, and a cam that meshes with the drive gear 405b provided on the camshaft 407b1
  • the method of constant-velocity rotational drive by the driven gear 406b provided on the shaft 407b2 is the same as that of the fourteenth embodiment, and thus the description thereof is omitted.
  • the centrifugal separation action of fuel having a density smaller than air, such as hydrogen supplied from the tangential port 230a of the internal combustion engine 1b and the injector 12a, is the same as in the first to third embodiments of the two-stroke internal combustion engine Do.
  • FIG. 33 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the fifteenth embodiment (FIG. 32).
  • a hydrogen premixed gas which is adiabatically compressed at TDC and has a temperature below the ignition point (300 ° C) of gasoline is spark-ignited by the spark plug 11b1 provided near the intersection of the combustion chamber and the cylinder axis.
  • the high hydrogen concentration layer (F1b) starts high-speed combustion, flame propagation at high temperature and high pressure expands uniformly in the circumferential direction from the center of the cylinder axis rotating at high speed by swirl and diffuses further to the middle concentration layer (F2b) Since hydrogen and gasoline also start combustion by the flame propagation, the pressure and temperature in the combustion chamber increase the hydrogen ignition point (585 ° C) and the hydrogen and gasoline such as low concentration layer (F3b) diffuses. Ignition of hydrogen in the region where it is located promotes combustion in the entire combustion chamber to suppress the generation of unburned gas.
  • the embodiments 1 to 15 describe an example of the present invention, and the internal combustion engine of each embodiment may be a diesel engine or a spark ignition internal combustion engine unless restricted, and the supply of fuel is not restricted.
  • the air flow amplifier may be an intake system or a combustion chamber, and the air flow amplifier of the supercharging amplification means may be any of an ejector, a flow transformer vector, a transformer vector, etc.
  • the compressor has been described as a reciprocating compressor.
  • the hybrid vehicles may be parallel or series.
  • the embodiments 1 to 15 show an example of the present invention and do not limit the present invention, and those skilled in the art can change and improve the present invention.
  • the internal combustion engine according to the present invention is a two-stroke internal combustion engine that does not require mixing of lubricating oil, has the same maintainability as a four-stroke internal combustion engine, and can perform reliable and satisfactory combustion with a simple configuration. Since sufficient scavenging and supercharging can be performed with a small displacement reciprocating compressor by simple scavenging amplification means, the improvement of scavenging improves the combustibility and the exhaust characteristics, and the output per unit volume of the internal combustion engine increases. Since the internal combustion engine can be downsized (small and lightweight), it can be used for an internal combustion engine mounted on a moving body such as a car or a ship. Also in the four-stroke internal combustion engine of claim 4, the internal combustion engine can be downsized and reduced in size and weight by the improvement of the combustion efficiency of the internal combustion engine.

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Abstract

[Problem] A conventional internal combustion engine has a problem in that it is impossible to efficiently and reliably collect a fuel at a spark plug in a combustion chamber and to form a combustible layer in which flame propagation becomes uniform in the circumferential direction, thus making it difficult to improve the combustion efficiency. [Solution] Provided is a two-cycle internal combustion engine in which: a combustion chamber is formed to have a substantially spherical surface; an intake valve and an exhaust valve are radially disposed in the combustion chamber; a spark plug or an injector is provided near a point of intersection with a cylinder axis of the combustion chamber; an intake port is formed of a tangential port; a fuel, such as hydrogen, having a smaller density than air is supplied; and a scavenging-air supply means that can supply scavenging air of a larger volume than a displacement volume is further provided, thereby making it possible to move the fuel, such as hydrogen, toward the cylinder axis through the action of centrifugation caused by a strong swirl continued until the end of a compression stroke and to perform stratified combustion. Provided is a two-cycle or four-cycle internal combustion engine that is provided with a combustion means in which the exhaust valve is opened and closed by cams provided on two camshafts, and the intake valve is opened and closed by a hydraulic mechanism.

Description

内燃機関Internal combustion engine
本発明は、２サイクル内燃機関の燃焼手段に関するものと、４サイクル内燃機関の燃焼手段と弁駆動機構に関するものである。 The present invention relates to the combustion means of a two-stroke internal combustion engine, and to the combustion means and valve drive mechanism of a four-stroke internal combustion engine.
２サイクル内燃機関（２ストローク1サイクルエンジン）は、回転数当たり４サイクル内燃機関の2倍の燃焼行程が発生するので大きな出力が得られるが、燃焼行程と圧縮行程との間に、排気を加圧された吸気でガス交換する掃気を行うので、燃焼行程と圧縮行程のストロークが縮減されるので、燃焼行程では膨張ストロークの減少による出力効率の低下、圧縮行程圧では圧縮ストロークの短縮による圧縮比の低下が発生する。
クランクケースを掃気の加圧ポンプとする場合は、潤滑のために吸気（または燃料）に潤滑油を混合する必要があり、吸排気方式に用いられるバルブレス・シリンダーポート方式はシリンダ側面の吸排気の専用孔（ポート）の開閉に往復運動するピストンの側面を利用するので、複雑な弁駆動機構が不要となる利点があるが、吸気に潤滑油を混入するので排気で潤滑油を放出して潤滑油を消費する問題点がある。
更に、吸気に混入する潤滑油は燃焼により煤や有害物質を発生するので排気性状が悪化する問題点がある。
吸気で排気をガス交換する掃気は、短絡掃気による吹き抜けが発生するので掃気効率が低下し、掃気にて完全なガス交換を行うにはシリンダ排気量より大きな容量の掃気が必要となり、過給を行う場合は、更に大容量の掃気供給手段が必要となる。 The two-stroke internal combustion engine (two-stroke one-stroke engine) generates twice as much combustion stroke as the four-stroke internal combustion engine per revolution, and thus a large output can be obtained. However, the exhaust is added between the combustion stroke and the compression stroke. Since the scavenging that performs gas exchange with compressed intake is performed, the strokes of the combustion stroke and the compression stroke are reduced, so the expansion stroke decreases in the combustion stroke, the output efficiency decreases by the compression stroke, and the compression ratio by the compression stroke decreases in the compression stroke pressure. Decline of
When using a crankcase as a scavenging pressurized pump, it is necessary to mix lubricating oil into the intake air (or fuel) for lubrication, and the valveless cylinder port system used for the intake and exhaust system is an Since the side of the reciprocating piston is used to open and close the dedicated hole (port), there is an advantage that a complicated valve drive mechanism is not required. However, since the lubricating oil is mixed in the intake, the lubricating oil is released by the exhaust to lubricate There is a problem of consuming oil.
Furthermore, since lubricating oil mixed in the intake air generates soot and harmful substances by combustion, there is a problem that the exhaust property is deteriorated.
Scavenging in which exhaust gas is exchanged by intake causes blowout by short circuit scavenging, so scavenging efficiency decreases, and in order to complete gas exchange by scavenging, scavenging with a larger capacity than cylinder displacement is required, and supercharging In the case of carrying out, a larger capacity scavenging gas supply means is required.
エンジン特性に対応して過給容量を設定するようにした場合には、過給機の容量を必要以上に大きくしなければならない場合があり、過給機が大型化するという問題点があり、排気ガスにより回転駆動される過給機を備えた内燃機関において、該過給機と燃焼室との間の吸気通路に、少なくとも１気筒のピストンストローク容積より大きい蓄圧室を設け、該蓄圧室の上流側に逆止弁を配置して、エンジン１回転の間の要求過給量の均一化を図ることができ、その分だけ過給機の容量を小さくでき、ひいては小型化に貢献できる過給機を備えた内燃機関（特許文献１）がある。
圧縮点火“スプリットサイクル”エンジンにて、ダスト放出や窒素酸化物の放出の少ない圧縮点火エンジンとし、更に、ターボスーパチャージャ（請求項１２）により過給ができるスプリットサイクルエンジン（特許文献２）がある。
これらの過給機（特許文献１、特許文献２）は、シリンダ排気量より大きな容量のポンプ能力を必要とするので、内燃機関の更なる小型化が困難となる問題点がある。 If the supercharging capacity is set according to the engine characteristics, the capacity of the supercharger may need to be increased more than necessary, and there is a problem that the supercharger becomes larger, In an internal combustion engine equipped with a turbocharger rotationally driven by exhaust gas, an accumulator chamber larger than a piston stroke volume of at least one cylinder is provided in an intake passage between the turbocharger and the combustion chamber, A check valve can be arranged on the upstream side to equalize the required supercharging amount during one rotation of the engine, and the capacity of the supercharger can be reduced by that amount, which can contribute to downsizing There is an internal combustion engine (Patent Document 1) equipped with a motor.
A compression-ignition "split-cycle" engine is a compression-ignition engine that emits less dust and releases nitrogen oxides, and there is also a split-cycle engine (Patent Document 2) that can be supercharged by a turbosupercharger (claim 12) .
Since these superchargers (patent documents 1 and 2) require a pump capacity of a capacity larger than the cylinder displacement, there is a problem that further miniaturization of the internal combustion engine becomes difficult.
内燃機関において、燃焼速度（火炎伝播速度）の向上を目的に、吸気流動である吸気スワール（以下「スワール）という。）を減衰させずに燃焼行程に移行するためには、球面状の燃焼室形状が望ましいが、従来の内燃機関ではエンジンに駆動されるカム機構の動作を機械式の伝達機構により弁に伝達して弁の開閉を行うので、燃焼室形状がペントルーフ形状に制約され、スワールには好ましくない形状に設計上制約される問題点がある。
この問題点の解決策として、前記機械式の伝動機構を流体圧式の伝動機構にすることにより、球面状の燃焼室と、前記燃焼室に放射状に配設される液体圧駆動式の弁と、エンジンにて駆動されるカム機構と、前記カム機構の動作を、流体圧を介して前記弁に伝達して前記弁を開閉させる流体機構とから成る内燃エンジン（特許文献３）がある。
燃焼室が球面状であるのでスワールの減衰を抑制できる効果と、燃焼室表面積が小さく冷却損失の抑制効果がある。
また、流体圧式の伝動機構として、４サイクル内燃機関で駆動する容積型ポンプを、ロータと、管状のカムと、ロータの回転軸に設けたベーンまたはプランジャとで構成し、カムを共用して多数の油圧回路を配置でき、簡素な構造により油圧供給手段の信頼性が高く、小型で安価に製作できる内燃機関の弁駆動機構（特許文献４）がある。 In an internal combustion engine, in order to shift to a combustion stroke without attenuating an intake swirl (hereinafter referred to as "swirl"), which is an intake flow, for the purpose of improving a combustion speed (flame propagation speed), a spherical combustion chamber Although the shape is desirable, in the conventional internal combustion engine, the operation of the cam mechanism driven by the engine is transmitted to the valve by the mechanical transmission mechanism to open and close the valve, so the shape of the combustion chamber is restricted to the pentroof shape, causing swirl. There is a problem that the design is constrained by the undesirable shape.
As a solution to this problem, by making the mechanical transmission mechanism a hydraulic transmission mechanism, a spherical combustion chamber and a liquid pressure drive valve radially disposed in the combustion chamber, There is an internal combustion engine (Patent Document 3) including a cam mechanism driven by an engine and a fluid mechanism that transmits the operation of the cam mechanism to the valve via fluid pressure to open and close the valve.
Since the combustion chamber is spherical, there is an effect of suppressing the attenuation of the swirl, and a reduction in the surface area of the combustion chamber and the cooling loss.
Also, as a hydraulic type transmission mechanism, a positive displacement pump driven by a four-cycle internal combustion engine is composed of a rotor, a tubular cam, and vanes or plungers provided on the rotary shaft of the rotor. There is a valve drive mechanism (patent document 4) of an internal combustion engine which can arrange the hydraulic circuit of the above and has a high reliability of the hydraulic pressure supply means by a simple structure, and can be made compact and inexpensive.
内燃機関の燃焼として、燃焼範囲（Ｖｏｌ％）が水素（４．１～７５）はガソリン（１～７．８）より広く、最小着火エネルギ（ｍｊ）が水素（０．０２）はガソリン（０．２４）より小さく、最大燃焼速度（ｃｍ／ｓ）が水素（３４６）はガソリン（４２）より大きいので、水素は点火しやすく、爆風圧が大きい利点があるが、燃料としては発熱量が小さく、エネルギ密度が小さい問題点がある。
水素等の気体燃料を有効に利用して燃料混合気のリーンリミットを拡大し、熱効率の向上と排気の改善とを実現させる目的で、燃料混合気を形成するための主燃料とは別の気体燃料を筒内に添加する内燃機関の気体燃料添加方法であって、前記気体燃料を吸気行程中に添加する第１ の添加行程と、前記気体燃料を圧縮行程中に添加する第２ の添加行程とを備える内燃機関の気体燃料添加方法（特許文献５）がある。
ノッキングの発生を効果的に抑制する目的で、ガソリン燃料を前記シリンダ内に噴射する筒内噴射ノズルと、水素燃料を吸気ポートにて噴射するポート噴射ノズルとを備える内燃機関で、一つの前記シリンダに対して一対の前記吸気ポートおよび一対の前記吸気バルブが設けられ、前記吸気バルブの開閉を制御する制御手段を備え、且つ、ノッキングが発生し易い運転領域にて、前記ポート噴射ノズルが前記吸気ポートにて水素燃料を噴射し、一つの前記シリンダに設けられた一対の前記吸気バルブのうち一方の前記吸気バルブが開弁すると共に他方の前記吸気バルブが閉弁して前記シリンダ内に吸気のスワールを発生させる内燃機関（特許文献６）がある。 The combustion range (Vol%) of hydrogen (4.1 to 75) is wider than gasoline (1 to 7.8), and the minimum ignition energy (mj) of hydrogen (0.02) is gasoline (0%) for combustion of internal combustion engines. .24) smaller and hydrogen (346) is larger than gasoline (42) because the maximum burning rate (cm / s) is larger than hydrogen (42), so it is easy to ignite and has a large blast pressure. There is a problem of low energy density.
A gas different from the main fuel for forming the fuel mixture for the purpose of expanding the lean limit of the fuel mixture by effectively using gaseous fuel such as hydrogen and realizing the improvement of the thermal efficiency and the improvement of the exhaust A method of gaseous fuel addition of an internal combustion engine, wherein fuel is added into a cylinder, a first addition step of adding the gaseous fuel during an intake stroke, and a second addition step of adding the gaseous fuel during a compression stroke And a gaseous fuel addition method for an internal combustion engine (Patent Document 5).
An internal combustion engine comprising an in-cylinder injection nozzle for injecting gasoline fuel into the cylinder and a port injection nozzle for injecting hydrogen fuel at an intake port for the purpose of effectively suppressing the occurrence of knocking; A pair of intake ports and a pair of intake valves are provided, control means for controlling the opening and closing of the intake valves is provided, and the port injection nozzle performs intake The hydrogen fuel is injected at the port, and one of the intake valves of the pair of intake valves provided in one of the cylinders is opened and the other of the intake valves is closed to perform intake of air in the cylinder. There is an internal combustion engine (patent document 6) which generates a swirl.
ディーゼル機関の燃焼として、発火点が水素（５００℃）はディーゼル燃料油（２２５℃）より高いので、断熱圧縮の着火時の温度を制御して水素を予混合するディーゼル機関等にてＨＣＣＩエンジン（予混合圧縮自動着火）またはＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）とすることにより燃焼性を改善できる。
広範囲の運転領域にわたり安定して運転できる予混合圧縮着火燃焼方式の内燃機関の制御装置として、軽油又は軽油を含む混合燃料をエンジンに供給する燃料供給系と、水素をエンジンに供給するガス供給系と、水素添加濃度によって変化する複数の燃焼波形を予めデータとして有し利用する要求予混合ガス演算部とを備え、エンジンの状態に応じて熱効率が高くなる様に、複数の燃焼波形の中から適切な一つを選択し、燃焼波形に一致する様にエンジンに供給する水素添加濃度を決定することにより、ＰＭ及びＮＯｘの生成量を低減できるとともに、エンジンの熱効率を向上できる内燃機関の制御装置（特許文献７）がある。
僅かな含水素ガス添加量で熱効率の向上やスート排出量の低減などを目的に、気筒に連通する複数の吸気ポートと、吸気に含水素ガスを添加する水素インジェクタと、気筒内に含軽油燃料を噴射する燃料インジェクタと、を備え、複数の吸気ポートは、ヘリカルポートであるセカンダリ吸気ポートとタンゼンシャルポートであるプライマリ吸気ポートを含み、上記水素インジェクタは、これら吸気ポートのうち、セカンダリ吸気ポートを介して気筒に導入される吸気にのみ含水素ガスを添加するディーゼル内燃機関（特許文献８）がある。 As combustion of a diesel engine, the ignition point is higher than hydrogen (500 ° C) than diesel fuel oil (225 ° C), so an HCCI engine (a diesel engine etc. that pre-mixes hydrogen by controlling the temperature at the time of adiabatic compression ignition) The flammability can be improved by using a premixed compression auto-ignition) or an SPCCI engine (compression-ignition combustion by spark control).
As a control device for a homogeneous charge compression ignition combustion type internal combustion engine capable of stably operating over a wide range of operation range, a fuel supply system for supplying light oil or a mixed fuel containing light oil to the engine, and a gas supply system for supplying hydrogen to the engine And a required premixed gas operation unit that has, as data, a plurality of combustion waveforms that change depending on the concentration of hydrogen addition, and the thermal efficiency is increased according to the state of the engine, from among the plurality of combustion waveforms A control device for an internal combustion engine capable of reducing the amount of PM and NOx produced and improving the thermal efficiency of the engine by selecting an appropriate one and determining the hydrogen addition concentration supplied to the engine so as to match the combustion waveform. (Patent Document 7).
For the purpose of improving thermal efficiency and reducing soot emissions with a small amount of hydrogen-containing gas added, multiple intake ports communicating with the cylinder, a hydrogen injector that adds hydrogen-containing gas to the intake, and light oil fuel And a plurality of intake ports include a secondary intake port which is a helical port and a primary intake port which is a tangential port, and the hydrogen injector is configured to set a secondary intake port among the intake ports. There is a diesel internal combustion engine (Patent Document 8) in which hydrogen-containing gas is added only to the intake air introduced into the cylinder via.
燃料の水素は燃料改質にて発生することもできるが、燃料を消費する問題点がある。
水素は水（電解液）を電気分解して水素と酸素を発生することもでき、電解液に超音波振動を伝搬し、電極から効率よく水素と酸素または酸水素を発生できる酸水素発生装置（特許文献９）があり、前記酸水素発生装置は水素、または酸水素を燃料とするハイブリッド車両等に搭載できる。 Although fuel hydrogen can be generated by fuel reforming, there is a problem of consuming fuel.
Hydrogen can also electrolyze water (electrolyte solution) to generate hydrogen and oxygen, propagate ultrasonic vibration to the electrolyte, and efficiently generate hydrogen and oxygen or oxygen from the electrode. Patent Document 9) can be mounted on a hybrid vehicle or the like using hydrogen or hydrogen oxide as a fuel.
少量の燃料を燃焼室の中心に分布させて点火性を向上し、周囲の吸気と十分に混合して燃焼効率を向上する、あるいは、水素は着火性がよく爆風圧が大きいので、先に水素雰囲気に点火して燃焼速度を向上して他の燃料の燃焼を促進する等の前記従来技術において、火花点火式、圧縮着火式に係らず、燃焼が始まる燃焼室の中央から燃焼性がよい燃料を層状に配置し、且つ、吸気の酸素を有効に活用するために燃料と均一に混合することが望まれる。
この望ましい状況を実現するために、燃焼室を球面状または略円錐状の軸対称形状とし、二つの吸気ポートから発生する強いスワールにて吸気と燃料を混合し、圧縮行程中に水素等の空気より密度が小さい燃料を遠心分離作用により点火プラグまたはインジェクタを配置した燃焼室のシリンダ軸近傍に集め、残余を燃料の濃度により層状に分離して成層燃焼により燃焼性を向上する４サイクル内燃機関（特許文献１０）がある。
前記４サイクル内燃機関では、吸入行程と圧縮行程での２ストローク間でスワールによる遠心分離作用が働くが、２サイクル内燃機関では、排気行程と圧縮行程の間の掃気行程と圧縮行程の１ストロークに満たない間となり、約半分のストロークとなるので前記遠心分離作用が十分に働かない。 A small amount of fuel is distributed in the center of the combustion chamber to improve ignition performance, and sufficiently mixed with the surrounding intake air to improve combustion efficiency, or hydrogen has good ignition performance and a large blast pressure. In the above-mentioned prior art, such as ignition of the atmosphere to improve the combustion rate to promote the combustion of other fuels, regardless of spark ignition type or compression ignition type, fuel having good combustibility from the center of the combustion chamber where combustion starts It is desirable to arrange the U.S. in layers and mix them uniformly with the fuel to make effective use of the oxygen in the intake.
In order to realize this desirable situation, the combustion chamber is formed into a spherical or substantially conical axisymmetric shape, and the strong swirl generated from the two intake ports mixes the intake and fuel, and air such as hydrogen during the compression stroke A four-cycle internal combustion engine that collects lower density fuel by centrifugal separation near the cylinder axis of the combustion chamber in which the spark plug or injector is disposed and separates the remainder into layers according to the concentration of fuel to improve flammability through stratified combustion ( Patent Document 10).
In the four-stroke internal combustion engine, the centrifugal separation action by swirl acts between two strokes in the intake stroke and the compression stroke, but in the two-stroke internal combustion engine, one stroke of the scavenging stroke between the exhaust stroke and the compression stroke and the compression stroke The centrifugal action does not work sufficiently because the stroke is less than half and the stroke is about half.
本発明の請求項１は、排気量より大きい容量の掃気を供給できる掃気供給手段を備えることにより、掃気の吹き抜けによる充填効率の低下を解消し、速度が減衰した初期流入吸気を排出し、掃気行程中に流れる掃気量が増大するので吸気流入速度が増大した強いスワールにより内燃機関の燃焼性を向上し、更に過給による出力増大効果によりダウンサイジングができる。
燃焼室形状を略球面状または略円錐状とし、燃焼室に放射状に吸気弁と排気弁を配置し、点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生するタンゼンシャルポートとし、前記掃気供給手段により、積極的に掃気の吹き抜けを行うことにより燃焼室に強いスワールを発生し、前記燃焼室形状によりスワールが圧縮行程のＴＤＣまで円滑に継続し、圧縮行程による燃焼室形状のドーム状への変形によりスワールの縮径が発生し、シリンダの中心部に生じる強いスワールの遠心分離作用により水素等の空気より密度の低い燃料を密度勾配に応じた層状に分離し、燃焼性がよい水素等を燃焼室のシリンダ軸中心に集めて成層燃焼により燃焼性が向上でき、更に火花点火式内燃機関またはディーゼル機関の燃料を供給する場合も燃焼性が向上できる。
シリンダヘッドに前記吸気弁と排気弁を設けるので燃料に潤滑油を混合する必要が無く、２サイクル内燃機関の問題点である排気に煤が発生する問題点が解消される。
更に、掃気供給手段として、内燃機関にて駆動する圧縮機と、吸気通路に空気流量増幅器を備えた簡素な構成の掃気増幅手段を設け、内燃機関の排気量以上の吸気を供給できるので、吸気の吹き抜け分を補充して確実なガス交換により燃焼性を向上し、排気弁閉鎖後に加圧掃気を供給して過給ができる。
本発明の請求項２は、複数の前記吸気弁と排気弁を交互に配置し、前記クランク軸により同期回転駆動する２本のカム軸を設け、前記２本のカム軸に前記排気弁を開閉するカムを各１個設け、前記吸気弁を逆止弁とする、または、カムで作動する油圧手段による前記吸気弁の駆動機構を設け、シリンダ軸対称に発生するスワールにより燃焼性を向上できる。本発明の請求項３は、電気的手段により運転する酸水素発生装置を設け、前記酸水素発生装置で発生する酸水素、または水素と酸素を前記内燃機関に供給し、水素より保管が容易な電解液貯蔵とし、電気的手段を利用したエネルギ回生手段にできる。 In the first aspect of the present invention, by providing the scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the displacement, the reduction of the filling efficiency due to the scavenging blowout is eliminated, and the initially inflowing intake air whose speed is attenuated is discharged. Since the amount of scavenging air flowing during the stroke is increased, the strong swirl with increased intake air inflow speed improves the combustibility of the internal combustion engine, and further downsizing can be performed by the power increase effect by supercharging.
The combustion chamber has a substantially spherical shape or a substantially conical shape, and an intake valve and an exhaust valve are radially arranged in the combustion chamber, an ignition plug or an injector is provided near the intersection with the cylinder axis of the combustion chamber, and an intake port is in the cylinder. The tangential port generates swirl, and the scavenging air supply means actively blows through scavenging to generate a strong swirl in the combustion chamber, and the shape of the combustion chamber smoothly causes the swirl to reach TDC in the compression stroke. Continued, the diameter of the swirl is reduced due to the deformation of the shape of the combustion chamber into a dome due to the compression stroke, and the centrifugal action of the strong swirl generated at the center of the cylinder causes a density gradient to be less dense fuel such as hydrogen. Can be separated into appropriate layers, and hydrogen and the like with good combustibility can be collected around the cylinder axis of the combustion chamber to improve the combustibility by stratified combustion, and further spark-ignition internal combustion When supplying fuel Seki or diesel engine can be improved combustibility.
Since the cylinder head is provided with the intake valve and the exhaust valve, there is no need to mix lubricating oil with fuel, and the problem of soot being generated in the exhaust which is a problem of the two-stroke internal combustion engine is solved.
Further, as the scavenging gas supply means, a compressor driven by the internal combustion engine and a scavenging gas amplification means having a simple configuration provided with an air flow rate amplifier in the intake passage can be provided to supply intake air larger than the displacement of the internal combustion engine. The blow-by portion is replenished to improve the combustibility by reliable gas exchange, and after the exhaust valve is closed, pressurized scavenging gas can be supplied to perform supercharging.
According to a second aspect of the present invention, a plurality of intake valves and exhaust valves are alternately arranged, and two camshafts rotationally driven synchronously by the crankshafts are provided, and the exhaust valves are opened and closed on the two camshafts. One cam can be provided, and the intake valve can be a check valve, or a drive mechanism of the intake valve by a hydraulic unit operated by the cam can be provided, and the combustibility can be improved by the swirl generated in the cylinder axis symmetry. According to a third aspect of the present invention, an oxyhydrogen generator operated by an electrical means is provided, and the oxyhydrogen generated by the oxyhydrogen generator, or hydrogen and oxygen are supplied to the internal combustion engine, and storage is easier than hydrogen. It can be used as electrolytic solution storage and energy regeneration means using electrical means.
本発明の請求項４は、４サイクル内燃機関の燃焼室を略球面状または略円錐状とし、燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、回転方向が同じスワールを発生する複数のタンゼンシャルポートを設け、吸気系統および／または燃焼室に水素、メタンのように空気より密度が小さい燃料等を供給する内燃機関において、更に、前記前記往復動機関のクランク軸の回転数の１／２の回転数で駆動される平行な２本のカム軸を設け、前記２本のカム軸に前記排気弁を開閉するカムを設け、前記カムに連動する油圧手段により開閉するまたは前記吸気弁を前記カムとは別のカムで作動する油圧手段により前記吸気弁を
開閉し、シリンダ軸対称に発生するスワールにより燃焼性を向上できる。 According to a fourth aspect of the present invention, the combustion chamber of the four-stroke internal combustion engine is substantially spherical or substantially conical, and a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber to generate swirls having the same rotational direction. In an internal combustion engine provided with a plurality of tangential ports for supplying hydrogen or fuel such as methane having a density smaller than air to an intake system and / or a combustion chamber, the rotation of the crankshaft of the reciprocating engine Two parallel camshafts driven at half the number of revolutions of the number are provided, and the two camshafts are provided with cams for opening and closing the exhaust valve, and are opened and closed by hydraulic means interlocking with the cams or The intake valve can be opened and closed by hydraulic means that operates the intake valve with a cam different from the cam, and the combustibility can be improved by the swirl generated in a cylinder axial symmetry.
特開平７－３１７５５５号公報Unexamined-Japanese-Patent No. 7-317555 特表２０１３－５０５３９６号公報Japanese Patent Application Publication No. 2013-505396 実開平０２－０９６４０３号公報Japanese Utility Model Publication No. 02-096403 特許第６１９０９９７号公報Patent No. 6190997 特開２００４－０７６６７９号公報Japanese Patent Application Publication No. 2004-076679 特開２０１０－２１６３９５号公報Unexamined-Japanese-Patent No. 2010-216395 特開２０１０－２５５４４２号公報JP, 2010-255442, A 特開２０１３－８３１９３号公報JP, 2013-83193, A 特許第６０９７９８７号公報Patent No. 6097987 特許第６２０９８０２号公報Patent No. 6209802
従来の内燃機関は、燃焼室の点火プラグ附近に燃料を効率よく確実に集めることができず、火炎伝播が周方向に均一になる可燃層を形成できないので燃焼効率の向上が困難である問題点がある。 In the conventional internal combustion engine, the fuel can not be collected efficiently and reliably near the spark plug in the combustion chamber, and the combustion efficiency can not be improved because the combustible layer which makes the flame propagation uniform in the circumferential direction can not be formed. There is.
請求項１～３は２サイクル内燃機関の燃焼手段に関するもので、請求項４は４サイクル内燃機関の燃焼手段に関するものである。 The first to third aspects relate to the combustion means of the two-stroke internal combustion engine, and the fourth aspect relates to the combustion means of the four-stroke internal combustion engine.
請求項１は、シリンダヘッドに吸気弁と排気弁を設けた２サイクル内燃機関において、排気量より大きい容量の掃気を供給できる掃気供給手段を備え、前記掃気供給手段は、前記内燃機関にて駆動する圧縮機と、吸気通路に掃気増幅手段を設け、前記掃気増幅手段は、逆止弁と前記逆止弁の下流に設けた空気流量増幅器から成り、前記空気流量増幅器の駆動流通路を前記圧縮機の吐出口に連通し、更に、燃焼室を略球面状または略円錐状とし、前記燃焼室に放射状に吸気弁と排気弁を配置し、点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポートとし、水素、メタンのように空気より密度が小さい燃料、または前記空気より密度が小さい燃料と火花点火式内燃機関またはディーゼル機関の燃料を吸気系統および/または前記燃焼室に供給し、前記内燃機関の運転状況に応じて前記燃料の供給を制御する２サイクル内燃機関である。 A first aspect of the present invention is a two-stroke internal combustion engine provided with an intake valve and an exhaust valve in a cylinder head, comprising: scavenging air supply means capable of supplying scavenging air having a capacity larger than the exhaust amount; the scavenging air supply means driven by the internal combustion engine And a scavenging amplification means provided in the intake passage, the scavenging amplification means comprising a check valve and an air flow amplifier provided downstream of the check valve, the drive flow path of the air flow amplifier being compressed Furthermore, the combustion chamber has a substantially spherical or substantially conical shape, and an intake valve and an exhaust valve are radially arranged in the combustion chamber, and an ignition plug or an injector is connected to the cylinder axis of the combustion chamber. In the vicinity of the intersection point, the intake port is a tangential port that generates swirl in the cylinder, and hydrogen, a fuel such as methane having a density smaller than that of air, or a fuel having a density smaller than that of air Fuel flowers ignition type internal combustion engine or diesel engine is supplied to the intake system and / or the combustion chamber, a two-cycle internal combustion engine for controlling the supply of the fuel in accordance with the operating condition of the internal combustion engine.
請求項２は、前記燃焼室に放射状に複数の前記吸気弁と排気弁を交互に配置し、クランク軸の回転数と同じ回転数で連動する平行な２本のカム軸を設け、前記排気弁を前記２本の各カム軸に設けたカムにより開閉し、前記吸気弁を逆止弁とする、または、前記吸気弁を前記カムに連動する油圧手段により開閉するまたは前記吸気弁を前記カムとは別のカムに連動する油圧手段により開閉する請求項１に記載の２サイクル内燃機関である。 A plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber, and two parallel camshafts interlocked at the same rotational speed as the rotational speed of the crankshaft are provided, and the exhaust valve Are opened and closed by a cam provided on each of the two camshafts, and the intake valve serves as a check valve, or the intake valve is opened and closed by hydraulic means interlocking with the cam, or 2. The two-stroke internal combustion engine according to claim 1, wherein the two-stroke internal combustion engine is opened and closed by hydraulic means interlocked with another cam.
請求項３は、前記内燃機関に電気的手段により運転する酸水素発生装置を設け、前記酸水素発生装置で発生する水素または酸水素を、前記空気より密度が小さい燃料として供給する請求項１または請求項２に記載の２サイクル内燃機関である。 According to a third aspect of the present invention, the internal combustion engine is provided with an oxyhydrogen generator operated by electrical means, and hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air. It is a two-stroke internal combustion engine according to claim 2.
請求項４は、４サイクル内燃機関において、燃焼室を略球面状または略円錐状とし、前記燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発
生させるタンゼンシャルポートとし、水素、メタンのように空気より密度が小さい燃料、または前記空気より密度が小さい燃料と火花点火式内燃機関またはディーゼル機関の燃料を吸気系統および/または前記燃焼室に供給し、前記内燃機関の運転状況に応じて前記燃料の供給を制御し、更に、クランク軸の回転数の１／２の回転数で連動する平行な２本のカム軸を設け、前記排気弁を前記２本の各カム軸に設けたカムにより開閉し、前記吸気弁を前記カムに連動する油圧手段により開閉するまたは前記吸気弁を前記カムとは別のカムに連動する油圧手段により開閉する４サイクル内燃機関である。 According to a fourth aspect of the present invention, in the four-stroke internal combustion engine, the combustion chamber is substantially spherical or substantially conical, a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber, and an ignition plug or injector is disposed in the combustion chamber. Near the intersection with the cylinder axis of the cylinder, the intake port is a tangential port that generates a swirl in the cylinder, and fuel such as hydrogen and methane has a density smaller than that of air, or fuel whose density is smaller than that of air and spark ignition Fuel of the internal combustion engine or diesel engine is supplied to the intake system and / or the combustion chamber, and the supply of the fuel is controlled according to the operating condition of the internal combustion engine, and Hydraulic means provided with two parallel camshafts interlocking at rotational speed, opening and closing the exhaust valve by cams provided at the two camshafts, and interlocking the intake valve with the cam It is a four-stroke internal combustion engine that opens and closes or opens and closes the intake valve by hydraulic means linked to a cam other than the cam.
本発明の請求項１は、シリンダヘッドに吸気弁と排気弁を設けてクランクケースをポンプ室としないので、吸気に潤滑油を混合する必要がなく排気に煤が発生する問題点が解消できる。
燃焼室を略球面状または略円錐状とし、燃焼室に放射状に吸気弁と排気弁を配置し、点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生するタンゼンシャルポートとし、水素、メタンのように空気より密度が小さい燃料等を気系統および/または前記燃焼室に供給することにより、燃焼室に発生する吸気スワールにより、前記水素等の空気より密度が小さい燃料を、遠心分離作用により点火プラグまたはインジェクタを設けたシリンダ軸側に移動し、図３と図４に示すように、燃焼室形状は巨視的にみると圧縮行程にて略円筒状から略球面ドーム状に変化するので、吸気の径方向の燃料の層状分布は周辺部から中心部に移動し、スワールの旋回径が運動エネルギを保持して縮径するので角速度が大きくなり回転数が増大し、サイクロン効果のように遠心力の増大により強い遠心分離作用が発生する燃焼手段は４ストローク内燃機関の特許文献１０と同じであるが、本願発明の２ストローク内燃機関では、前記遠心分離作用を４ストローク内燃機関の約半分の１ストロークで行うので十分に行えない問題点がる。
そこで、排気量より大きい容量の掃気を供給できる掃気供給手段を設け、掃気行程で供給する掃気量を大きくすることにより掃気流入速度を増大し、排気との衝突によりスワールの速度が低下した流入初期の掃気流を吹き抜けによる短絡掃気として放出して強いスワールとし、前記燃焼室形状により圧縮行程終了時までスワールの回転運動を阻害されることなく前記強いスワールが継続できる。
前記強いスワールによる遠心分離作用により、気体の密度勾配に応じて前記水素等の燃料は図５に示すように層状に分離し、中心部に高濃度可燃層を形成し、燃焼室のシリンダ軸との交点近傍に設けた前記点火プラグまたはインジェクタの発火部に集まる。
前記遠心分離作用によりシリンダ軸側に図６に示すように密度（分子量）の小さい水素等の高濃度層を形成し、その周辺に水素より密度が大きい主燃料の可燃層を形成して、水素の確実で速い火炎伝播により前記主燃料の燃焼を促進する効果があり、ディーゼル機関では回転数の増大が可能となり出力増大効果がある。
水素等の前記高濃度可燃層にて確実に点火または着火ができ、火炎伝播は周方向に均等に順次高濃度層側から低濃度層に伝播し、最外側の超低濃度層では燃焼温度が低いので火炎伝播による燃焼室壁面の温度上昇が抑制されるので冷却損失を抑制し、前記成層燃焼により燃焼性が向上する効果との相乗効果により内燃機関の熱効率が増大する。
火炎伝播の到達する以前に急激に残りの未燃混合気が燃焼するノッキング現象は、密度勾配に応じた前記成層燃焼（最外層の超低濃度層）によりノッキング抑制効果がある。
水素を燃料とする場合は燃焼により水が発生し、少量の水または水蒸気は燃焼を助ける効果があるので排気性状が改善し、水は水蒸気になると約１７００倍に膨張して内燃機関の筒内圧力を増大するので出力増大効果がある。 In the first aspect of the present invention, since the intake valve and the exhaust valve are provided in the cylinder head and the crankcase is not used as the pump chamber, it is not necessary to mix lubricating oil into the intake, and the problem of generating soot in the exhaust can be solved.
The combustion chamber has a substantially spherical or substantially conical shape, the intake valve and the exhaust valve are radially arranged in the combustion chamber, the spark plug or the injector is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is in the cylinder. It is a tangential port that generates swirl, and hydrogen and other fuels with a density smaller than air such as methane are supplied to the air system and / or the combustion chamber, so that the hydrogen etc. is generated by the intake swirl generated in the combustion chamber. The fuel whose density is smaller than that of the air is moved to the cylinder shaft side provided with the spark plug or injector by centrifugal separation, and as shown in FIGS. 3 and 4, the shape of the combustion chamber macroscopically Since it changes from a substantially cylindrical shape to a substantially spherical dome shape, the laminar distribution of fuel in the radial direction of intake moves from the peripheral portion to the central portion, and the swirling diameter of the swirl holds kinetic energy. Because the diameter decreases, the angular velocity increases and the rotational speed increases, and the combustion means that produces a strong centrifugal separation effect due to the increase of centrifugal force like the cyclone effect is the same as Patent Document 10 of the four-stroke internal combustion engine In the two-stroke internal combustion engine of the present invention, the centrifugal separation operation is performed in about one half stroke of the four-stroke internal combustion engine, which is not sufficient.
Therefore, a scavenging air supply means capable of supplying scavenging air with a capacity larger than the exhaust gas volume is provided, and the scavenging air inflow speed is increased by increasing the scavenging air amount supplied in the scavenging stroke. The scavenging air flow is discharged as a short circuit scavenging air due to the blowout to make a strong swirl, and the strong swirl can be continued without the rotational movement of the swirl being inhibited by the shape of the combustion chamber until the end of the compression stroke.
Due to the centrifugal action by the strong swirl, the fuel such as hydrogen is separated into layers as shown in FIG. 5 according to the density gradient of the gas, forming a high concentration combustible layer in the center, and the cylinder axis of the combustion chamber Collecting at the ignition part of the spark plug or the injector provided near the intersection of
As shown in FIG. 6, a high concentration layer of hydrogen or the like having a small density (molecular weight) is formed on the cylinder axis side by the centrifugal separation action, and a combustible layer of the main fuel having a density larger than hydrogen is formed around it. In the diesel engine, the rotation speed can be increased and the output can be increased.
The high concentration combustible layer such as hydrogen can be ignited or ignited with certainty, flame propagation propagates uniformly in the circumferential direction sequentially from the high concentration layer side to the low concentration layer, and in the outermost ultra low concentration layer the combustion temperature is Since the temperature rise of the combustion chamber wall surface by flame propagation is suppressed since it is low, the cooling loss is suppressed, and the thermal efficiency of the internal combustion engine is increased by the synergistic effect with the effect of improving the combustibility by the stratified combustion.
The knocking phenomenon in which the remaining unburned air-fuel mixture burns rapidly before reaching the flame propagation has a knocking suppression effect due to the stratified combustion (ultra-low concentration layer in the outermost layer) according to the density gradient.
When using hydrogen as fuel, water is generated by combustion, and a small amount of water or steam has the effect of assisting combustion, so the exhaust properties are improved, and when water turns into steam, it expands about 1700 times and is in the cylinder of the internal combustion engine Since the pressure is increased, there is an output increase effect.
更に、圧縮機で発生する圧縮空気を駆動流とする空気流量増幅器から成る掃気増幅手段により、簡素で小さな容量の圧縮機で内燃機関の排気量以上の掃気を供給できる効果がある。前記掃気増幅手段は吸気通路の下流側に逆止弁を設け、空気流量増幅器の逆流量増幅現象（下流側の圧力上昇により流れが反転し、上流側に流量増幅を伴って逆流する現象）を防止し、逆止弁により高圧の駆動流を直接シリンダに流入して過給効果が発生する。
２サイクル内燃機関の掃気を、前記大容量の吸気により吹き抜けを伴っても十分に排気を排出できるので、完全なガス交換ができるので燃焼性の向上による出力増大効果と、前記大容量の吸気による吹き抜けにより熱負荷が大きい２サイクル内燃機関の燃焼室の冷却効果がある。
駆動流通路に燃料を掃気タイミング後半に供給し、過給と同時に駆動流により前記燃料を吸気に均一に予混合できるので、吸気中の酸素を効率よく燃焼に利用できる。
図２０に示すように、燃焼室の温度制御等によりＨＣＣＩエンジン（予混合圧縮自動着火）またはＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）とすることにより燃焼性の改善と出力の向上の効果がある。 Furthermore, scavenging amplification means comprising an air flow amplifier having a compressed air generated by the compressor as a drive flow has an effect of being able to supply scavenging air exceeding the displacement of the internal combustion engine with a simple and small-capacity compressor. The scavenging air amplification means is provided with a check valve on the downstream side of the intake passage, and a reverse flow amplification phenomenon of the air flow amplifier (a flow is reversed due to the pressure increase on the downstream side and a backflow phenomenon with flow amplification on the upstream side) The high pressure drive flow directly flows into the cylinder by the check valve to generate the supercharging effect.
Since scavenging of a two-stroke internal combustion engine can be sufficiently exhausted even with blowout by the large capacity intake, complete gas exchange can be performed, so the power increase effect by the improvement of the combustibility and the large capacity intake The blow through has a cooling effect on the combustion chamber of a two-stroke internal combustion engine having a large heat load.
Since fuel can be supplied to the drive flow passage in the latter half of the scavenging timing and the fuel can be uniformly premixed in the intake air by the drive flow simultaneously with supercharging, oxygen in the intake air can be efficiently used for combustion.
As shown in FIG. 20, the HCCI engine (premixed compression auto-ignition) or SPCCI engine (compression-ignition combustion by spark control) with the temperature control of the combustion chamber has the effect of improving the combustibility and the output. .
本発明の請求項２は、複数の前記吸気弁と排気弁を交互に配置し、前記クランク軸により同期回転駆動する２本のカム軸を設け、前記２本のカム軸に前記排気弁を開閉するカムを各１個設け、前記吸気弁を逆止弁とする、またはカムで作動する油圧手段による前記吸気弁の駆動機構を設け、シリンダ軸対称に発生するスワールにより燃焼性を向上でき、弁の開弁時に排気圧が作用する排気弁をカム駆動とすることにより高速高負荷運転時の信頼性が向上する。 According to a second aspect of the present invention, a plurality of intake valves and exhaust valves are alternately arranged, and two camshafts rotationally driven synchronously by the crankshafts are provided, and the exhaust valves are opened and closed on the two camshafts. The intake valve is provided as a check valve, or a drive mechanism for the intake valve is provided by hydraulic means operated by the cam, and the combustibility can be improved by the swirl generated in the cylinder axis symmetry, and the valve By driving the exhaust valve to which the exhaust pressure acts at the time of opening the valve as a cam drive, the reliability in high speed and high load operation is improved.
本発明の請求項３は、前記内燃機関に電気的手段により運転する酸水素発生装置を設け、前記酸水素発生装置で発生する水素または酸水素を、前記空気より密度が小さい燃料として供給することにより、水素の自給または補充量を減少でき、前記酸水素発生装置を運転する電気的手段に回生エネルギを利用し、２次電池の電気容量を抑制できる効果がある。 According to a third aspect of the present invention, the internal combustion engine is provided with an oxyhydrogen generator operated by electrical means, and hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air. As a result, the amount of self-supply or replenishment of hydrogen can be reduced, and regenerative energy can be used for the electric means for operating the oxyhydrogen generator to suppress the electric capacity of the secondary battery.
本発明の請求項４は、４サイクル内燃機関の燃焼室を略球面状または略円錐状とし、燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、クランク軸の回転数の１／２の回転数で駆動される２本のカム軸を設け、前記２本のカム軸に前記排気弁を開閉するカムを各１個設け、前記吸気弁をカムで作動する油圧手段による前記吸気弁の駆動機構を設けるので、シリンダ軸に対向配置したタンゼンシャルポートにより発生する強いスワールにより燃焼性を向上でき、開弁時に排気圧が作用する排気弁をカム駆動とすることにより高速高負荷運転時の信頼性が向上する。
４サイクルと２サイクルの違いがあるが、前記請求項１と請求項２と同様の作用と効果がある。 According to a fourth aspect of the present invention, the combustion chamber of the four-stroke internal combustion engine is substantially spherical or substantially conical, and a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber. Provided with two camshafts driven at a rotational speed of 1/2, each cam provided with one cam for opening and closing the exhaust valve, and the intake by hydraulic means operating the intake valve with the cam Since the drive mechanism of the valve is provided, the strong swirl generated by the tangential port arranged opposite to the cylinder axis can improve the combustibility, and the exhaust valve acting on the exhaust pressure at the time of valve opening is cam driven, high speed and high load Reliability during driving is improved.
Although there is a difference between 4 cycles and 2 cycles, the same actions and effects as in the above-mentioned claim 1 and claim 2 can be obtained.
実施例１（請求項１対応）の、略球面状の燃焼室に放射状に配置した吸気弁と排気弁、およびタンゼンシャル吸気ポートを設け、水素燃料を供給する２サイクル内燃機関の構成概念の説明図である。Explanatory drawing of the construction concept of a two-stroke internal combustion engine provided with an intake valve and an exhaust valve radially disposed in a substantially spherical combustion chamber and a tangential intake port according to Embodiment 1 (corresponding to claim 1) to supply hydrogen fuel. It is. 実施例２（請求項１対応）の、略球面状の燃焼室にタンゼンシャルポートを設けた水素とガソリンを燃料とする内燃機関の平面図と周辺回路図である。FIG. 7 is a plan view and a peripheral circuit diagram of an internal combustion engine using a hydrogen and a gasoline as fuel, in which tangential ports are provided in a substantially spherical combustion chamber according to Embodiment 2 (corresponding to claim 1). 前記実施例２（図２）の燃焼室の、圧縮行程の排気終了後の掃気中（Ｐ１）と圧縮行程の掃気終了後（Ｐ２）の各掃気挙動（スワール）の説明図である。It is explanatory drawing of each scavenging behavior (swirl) during scavenging (P1) after the completion | finish of exhaust_gas | exhaustion of the compression stroke of the combustion chamber of the said Example 2 (FIG. 2), and scavenging completion (P2) of a compression stroke. 前記実施例２（図２）の燃焼室の排気終了時（Ｕ１）と圧縮終了時（Ｕ２）の容積図、と圧縮行程の試算による径方向容積占有率の変化の説明図（Ｕ３）である。It is explanatory drawing (U3) of the change of the radial direction volume occupancy by the calculation of the volume at the completion | finish (U1) of the exhaust of a combustion chamber of the said Example 2 (FIG. 2), and completion | finish of compression (U2), and a compression stroke. . 前記実施例２（図２）の、ＴＤＣの水素の各燃料濃度層の分布状況と点火時の高速火炎伝播の説明図である。It is explanatory drawing of the distribution condition of each fuel concentration layer of the hydrogen of TDC of the said Example 2 (FIG. 2), and the high-speed flame propagation at the time of ignition. 前記実施例２（図２）の内燃機関等の燃料と空気の特性図で、燃料は燃焼範囲と密度、空気は組成割合と密度を示す。FIG. 2 is a characteristic diagram of fuel and air of the internal combustion engine etc. of the second embodiment (FIG. 2), wherein the fuel shows the combustion range and the density, and the air shows the composition ratio and the density. 実施例３（請求項１対応）の往復圧縮機、吸排気の弁と通路の配置を示す３気筒内燃機関の平面図と、掃気増幅手段と容積型油圧供給手段等の周辺回路図である。FIG. 14 is a plan view of a three-cylinder internal combustion engine showing the arrangement of a reciprocating compressor, intake and exhaust valves and passages according to a third embodiment (corresponding to claim 1), and peripheral circuits such as scavenging air amplification means and positive displacement hydraulic pressure supply means. 前記実施例３（図７）のＪ－Ｊ断面の吸気弁と排気弁の冷却手段を設けた前記内燃機関の断面図である。FIG. 8 is a cross-sectional view of the internal combustion engine provided with cooling means for the intake valve and the exhaust valve of the JJ cross section of the third embodiment (FIG. 7). 実施例４（請求項１対応）の逆止弁と空気流量増幅器から成る掃気増幅手段と、クランク軸より位相が進んだ往復圧縮機とを備えた２サイクル内燃機関の構成概念の説明図である。It is explanatory drawing of the construction concept of the 2 cycle internal combustion engine provided with the scavenging-gas amplification means which consists of a non-return valve and the air flow amplifier of Example 4 (Claim 1 claim), and a reciprocating compressor whose phase advanced from the crankshaft. . 前記実施例４（図９）の掃気増幅手段の構成例で、空気流量増幅器の流量増幅比の小さい順にエジェクタ（Ａ）、従来技術のフロートランスベクタ（Ｂ）とトランスベクタ（Ｃ）と逆止弁の構成説明図である。In the configuration example of the scavenging air amplification means of the fourth embodiment (FIG. 9), the ejector (A) and the flow transvector (B) and the transvector (C) of the prior art are reversed in ascending order of flow amplification ratio of the air flow amplifier. It is structure explanatory drawing of a valve. 前記実施例４（図９）の内燃機関の、排気行程初期（Ｓ１）、掃気行程（Ｓ２）、および圧縮行程（Ｓ３）の掃気増幅手段の動作説明図である。FIG. 14 is an operation explanatory view of scavenging amplification means of the exhaust stroke initial (S1), scavenging stroke (S2), and compression stroke (S3) of the internal combustion engine of the fourth embodiment (FIG. 9). 前記実施例４（図９）の内燃機関のタイミングダイアグラムである。It is a timing diagram of the internal combustion engine of the said Example 4 (FIG. 9). 前記実施例４（図９）の往復圧縮機と掃気増幅手段を設けた内燃機関の各部のタイミングチャートと試算による筒内圧力である。It is a cylinder pressure by trial calculation and the timing chart of each part of the internal combustion engine which provided the reciprocating compressor and scavenging air amplification means of said Example 4 (FIG. 9). 前記実施例４（図９）の掃気増幅手段を設けた内燃機関の高速回転時の試算によるＰＶ線図である。It is a PV diagram by trial calculation at the time of high speed rotation of an internal combustion engine provided with the scavenging air amplification means of the fourth embodiment (FIG. 9). 実施例５（請求項１対応）の、リフト逆止弁と従来技術のトランスベクタを設けた掃気増幅手段と往復圧縮機を設けた内燃機関の断面図である。It is sectional drawing of the internal combustion engine which provided the scavenging air amplification means which provided the lift non-return valve and the prior art trans | transformer vector, and the reciprocating compressor of Example 5 (corresponding to Claim 1). 前記実施例５（図１５）の従来技術の可変ノズル型のトランスベクタとリフト逆止弁で構成される掃気増幅手段の断面図である。It is sectional drawing of the scavenging air amplification means comprised with the variable nozzle type transvector and lift check valve of the prior art of the said Example 5 (FIG. 15). 実施例６（請求項１対応）の掃気増幅手段と往復圧縮機を設けた２気筒内燃機関１ｋの構成説明図の平面図（１）、Ｋ－Ｋ断面図（２）、Ｌ－Ｌ断面図（３）である。A plan view (1), a KK sectional view (2), and an LL sectional view of a configuration explanatory view of a two-cylinder internal combustion engine 1k provided with the scavenging air amplification means and the reciprocating compressor of the sixth embodiment (corresponding to claim 1) (3). 前記実施例６（図１７）の内燃機関の各気筒のタイミングチャートと試算による筒内圧力である。They are the in-cylinder pressure by trial calculation and the timing chart of each cylinder of the internal combustion engine of the said Example 6 (FIG. 17). 実施例７（請求項１対応）の掃気増幅手段と往復圧縮機を設けた３気筒内燃機関の構成説明図の平面図（１）とＭ－Ｍ断面図（２）である。FIG. 20 is a plan view (1) and an MM cross-sectional view (2) of a configuration explanatory view of a three-cylinder internal combustion engine provided with the scavenging air amplification means of the seventh embodiment (corresponding to claim 1) and a reciprocating compressor. 実施例８（請求項１対応）の往復圧縮機と、掃気増幅手段に制御弁を設けた２サイクル内燃機関の構成概念の説明図である。FIG. 18 is an explanatory view of a configuration concept of a two-stroke internal combustion engine in which a reciprocating compressor according to an eighth embodiment (corresponding to claim 1) and a control valve are provided to a scavenging air amplification means. 前記実施例８（図２０）の掃気増幅手段に設けた制御弁による掃気の流量増幅抑制時の試算によるＰＶ線図である。FIG. 21 is a PV diagram based on a trial calculation at the time of flow rate amplification suppression of scavenging air by a control valve provided in the scavenging air amplification means of the eighth embodiment (FIG. 20). 実施例９（請求項２対応）の燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、排気弁を２本の各カム軸に設けたカムにより開閉し、吸気弁を逆止弁とする掃気増幅手段を設けた２気筒２サイクル内燃機関の構成概念の説明図である。A plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber of the ninth embodiment (corresponding to claim 2), the exhaust valves are opened and closed by cams provided on each of two camshafts, and the intake valves are non-returned. FIG. 2 is an explanatory view of a configuration concept of a two-cylinder two-stroke internal combustion engine provided with a scavenging air amplification means as a valve. 実施例１０（請求項２対応）の２本のカム軸により、排気弁がカム駆動で吸気弁が油圧手段を介して油圧駆動する２気筒２サイクル内燃機関の平面図と周辺回路図である。FIG. 24 is a plan view and a peripheral circuit diagram of a two-cylinder two-stroke internal combustion engine in which an exhaust valve is cam-driven and an intake valve is hydraulically driven via hydraulic means by two camshafts of Embodiment 10 (corresponding to claim 2). 前記実施例１０（図２３）の、ＴＤＣの各燃料濃度層の分布状況と水素可燃層の点火時のＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）の説明図である。FIG. 24 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the tenth embodiment (FIG. 23). 実施例１１（請求項３対応）の掃気増幅手段と酸水素発生装置を備えた水素を燃料とする内燃機関の構成概念の説明図である。It is explanatory drawing of the structural concept of the internal combustion engine which uses hydrogen as a fuel provided with the scavenging air amplification means and the oxyhydrogen generator of Example 11 (corresponding to claim 3). 実施例１２（請求項３対応）の、従来技術の電解液に超音波を付加する酸水素発生装置の構成概念の説明図で、前記実施例１１（図２）の前記酸水素発生装置の一例である。FIG. 16 is an explanatory diagram of a conceptual configuration of an oxyhydrogen generator for adding an ultrasonic wave to a prior art electrolyte solution of Example 12 (corresponding to claim 3), which is an example of the oxyhydrogen generator of Example 11 (FIG. 2). It is. 実施例１３（請求項３対応）の掃気増幅手段、酸水素発生装置、と回生手段を設けたハイブリッド車両の火花点火式内燃機関の構成概念の説明図である。It is explanatory drawing of the construction concept of the spark-ignition internal combustion engine of the hybrid vehicle provided with the scavenging air amplification means, oxyhydrogen generator and regeneration means of Example 13 (corresponding to claim 3). 前記実施例１３（図２６）の前記ハイブリッド車両の内燃機関で、ＨＣＣＩエンジンまたはＳＰＣＣＩエンジンとして運転できる制御システムの構成概念の説明図である。FIG. 26 is an explanatory diagram of a configuration concept of a control system which can be operated as an HCCI engine or an SPCCI engine in the internal combustion engine of the hybrid vehicle in the thirteenth embodiment (FIG. 26). 前記実施例１３の内燃機関１ｔを組み込んだ前記ハイブリッド車両の内燃機関の制御システムをＨＣＣＩエンジンまたはＳＰＣＣＩエンジンとして運転する制御フローチャートである。FIG. 33 is a control flowchart for operating the control system of the internal combustion engine of the hybrid vehicle incorporating the internal combustion engine 1t of the thirteenth embodiment as an HCCI engine or an SPCCI engine. 実施例１４請求項４対応）の、略球形の燃焼室に吸気弁と排気弁を放射状に配置し、２本のカム軸の各カムで各排気弁の開閉と油圧手段を介して各吸気弁を開閉する４サイクル内燃機関の構成概念の説明図である。The intake valve and the exhaust valve are radially arranged in the substantially spherical combustion chamber of the fourteenth embodiment (corresponding to claim 4), and the respective cams of the two camshafts open and close the respective exhaust valves and the respective intake valves via hydraulic means. It is explanatory drawing of the structural concept of 4 cycle internal combustion engine which opens and closes. 前記実施例１４（図３０）の、ＴＤＣの水素の各燃料濃度層の分布状況と燃料噴射時のＨＣＣＩエンジンの燃焼の説明図である。FIG. 31 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and combustion of an HCCI engine at the time of fuel injection in the fourteenth embodiment (FIG. 30). 実施例１５請求項４対応）の、２本のカム軸に設けた各カムにより排気弁を開閉し、前記カムとは別のカムにより油圧手段を介して吸気弁を開閉する４サイクル内燃機関の平面図と周辺回路図である。Embodiment 15: A four-stroke internal combustion engine which opens and closes an exhaust valve by means of cams provided on two camshafts, and opens and closes an intake valve through hydraulic means by means of a cam different from the cam. It is a top view and a peripheral circuit diagram. 前記実施例１５（図３２）の、ＴＤＣの各燃料濃度層の分布状況と水素可燃層の点火時のＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）の説明図である。FIG. 32 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the fifteenth embodiment (FIG. 32).
前記図面（図１～３３）に従って、本願発明の各実施例（実施例１～１５）を、以下に説明する。
以下に説明する実施例は、制約が無い限り内燃機関は火花点火式内燃機関でもディーゼル機関でもよく、燃料供給は吸気系統および／または燃焼室でもよく、容積型油圧供給手段の油圧ポンプはベーンポンプでもプランジャーポンプでもよい。 Each example (Examples 1 to 15) of the present invention will be described below according to the above drawings (FIGS. 1 to 33).
In the embodiments described below, the internal combustion engine may be a spark ignition internal combustion engine or a diesel engine, without limitation, the fuel supply may be an intake system and / or a combustion chamber, and the hydraulic pump of the positive displacement hydraulic supply means may be a vane pump It may be a plunger pump.
図１は、実施例１（請求項１対応）の、略球面状の燃焼室に放射状に配置した吸気弁と排気弁、およびタンゼンシャル吸気ポートを設け、水素燃料を供給する２サイクル内燃機関の構成概念の説明図である。
図１は、シリンダヘッドに吸気弁４６と排気弁４７を設けた２サイクル内燃機関１において、排気量より大きい容量の掃気を供給できる掃気供給手段を備え、前記掃気供給手段は、前記内燃機関１にて駆動する圧縮機２５と、吸気通路である吸気流入通路２２と吸気流出通路２３の間に掃気増幅手段５を設け、前記掃気増幅手段５は、逆止弁５５と前記逆止弁５５の下流に設けた空気流量増幅器５０から成り、前記空気流量増幅器５０の駆動流通路５８を前記圧縮機２５の吐出口に連通し、更に、燃焼室を半径ＳＲの略球面状とし、前記燃焼室に放射状に吸気弁４６と排気弁４７を配置し、点火プラグ１１を前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポート２３０とし、水素のように空気より密度が小さい燃料を前記燃焼室に供給し、内燃機関１の運転状況に応じて前記燃料の供給を制御する２サイクル内燃機関１である。
内燃機関１は、内燃機関１の排気量以上の吐出量の往復圧縮機２５を設けたスプリットサイクルの２サイクル内燃機関１で、出力手段４の連結棒４３と同じ位相のクランク軸に前記往復圧縮機２５の連結棒２５３を設け、前記往復圧縮機２５のシリンダ２５１の直径（φＣ）を出力手段４のシリンダ４１の直径（φＥ）より小さくし、前記空気流量増幅器５０の流量増幅作用により排気量以上の容量の掃気を供給し、燃料タンク７５に貯蔵した水素を高圧燃料ポンプ１３で加圧してインジェクタ１２にて燃焼室に適時供給する。
吸気弁４６は弁駆動機構が不要な逆止弁とし、排気弁４７は、弁シリンダ４７１、弁ピストン４７２から成る油圧式アクチェータを油圧通路８８から供給される容積型油圧供給手段８で発生する油圧にて作動して弁を開閉する。
前記容積型油圧供給手段８は、クランク軸４４に同期回転するロータ８２と、カム８１、ベーン８３から成る油圧ポンプを備え、ロータ８２のベーン８３を増設することにより後述する実施例３（図７）のように多気筒に対応できる従来技術（特許文献４）の内燃機関の弁駆動機構である。 FIG. 1 shows the configuration of a two-stroke internal combustion engine provided with hydrogen fuel by providing an intake valve and an exhaust valve radially disposed in a substantially spherical combustion chamber and a tangential intake port according to Embodiment 1 (corresponding to claim 1). It is explanatory drawing of a concept.
FIG. 1 shows a two-stroke internal combustion engine 1 in which an intake valve 46 and an exhaust valve 47 are provided in a cylinder head, provided with scavenging air supply means capable of supplying scavenging air of a larger capacity than the exhaust gas. The scavenging amplification means 5 is provided between the compressor 25 driven at the same time and the intake inflow passage 22 and the intake and outflow passage 23 which are intake passages, and the scavenging amplification means 5 comprises a check valve 55 and a check valve 55. The air flow amplifier 50 is provided downstream, and the drive flow passage 58 of the air flow amplifier 50 is in communication with the discharge port of the compressor 25, and the combustion chamber has a substantially spherical shape with a radius SR. The intake valve 46 and the exhaust valve 47 are arranged radially, the spark plug 11 is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is a tangential port 230 that generates swirl in the cylinder. The fuel density than air is small is supplied to the combustion chamber as a 2-cycle internal combustion engine 1 for controlling the supply of the fuel in accordance with the operating condition of the internal combustion engine 1.
The internal combustion engine 1 is a split-cycle two-stroke internal combustion engine 1 provided with a reciprocating compressor 25 having a discharge amount equal to or greater than the displacement of the internal combustion engine 1, and the reciprocating compression is performed on the crankshaft having the same phase as the connecting rod 43 of the output means 4. The connecting rod 253 of the machine 25 is provided, and the diameter (φC) of the cylinder 251 of the reciprocating compressor 25 is made smaller than the diameter (φE) of the cylinder 41 of the output means 4. The scavenging air of the above capacity is supplied, and the hydrogen stored in the fuel tank 75 is pressurized by the high pressure fuel pump 13 and is supplied to the combustion chamber by the injector 12 as needed.
The intake valve 46 is a check valve that does not require a valve drive mechanism, and the exhaust valve 47 is a hydraulic pressure generated by the positive hydraulic pressure supply means 8 supplied from the hydraulic pressure passage 88 with a hydraulic actuator consisting of a valve cylinder 471 and a valve piston 472 Act to open and close the valve.
The positive displacement hydraulic pressure supply means 8 includes a hydraulic pump consisting of a rotor 82 rotating in synchronization with the crankshaft 44, a cam 81 and a vane 83, and a vane 83 of the rotor 82 is added to a third embodiment (FIG. The valve drive mechanism of an internal combustion engine according to the prior art (Patent Document 4) that can handle multiple cylinders as in the above.
図１の内燃機関１の作用は、分割サイクルを行う往復圧縮機２５と出力手段４から成る２サイクル内燃機関１は、前記往復圧縮機２５で内燃機関１の排気量以上の掃気を、前記タンゼンシャルポート２３０からシリンダ４１に供給してスワールを発生し、掃気行程で供給する掃気量を大きくすることにより掃気流入速度を増大し、排気との衝突によりスワールの速度が低下した流入初期の掃気流を吹き抜けによる短絡掃気として放出して強いスワールとし、前記燃焼室形状により圧縮行程終了時までスワールの回転運動を阻害されることなく前記強いスワールが継続できる。
内燃機関１は、シリンダヘッドに吸気弁と排気弁を設けた２サイクル内燃機関であるので、掃気効率が高いユニフロー掃気方式より低い掃気効率となるが、前記積極的な掃気の吹き抜けにより前記掃気効率を向上して、燃焼効率を向上できる効果がある。
排気弁４７の開閉は、容積型油圧供給手段８のベーン８３の作用角θｖとカム８１のカムプロフィール８８１の作用角θｃの各々の作用角の和が、油圧による排気弁４７の開弁タイミング角となる。
排気弁４７を開弁する油圧の油圧通路８８に連通する油圧補助手段８７の逆止弁８７５は、油圧油の漏れが発生した場合の油補充を行うので排気弁４７のラッシュアジャスタ作用がある。
排気弁４７の弁駆動機構は、弁配置が可能であれば他の駆動機構でもよい。
内燃機関１は、シリンダヘッドに吸気弁４６と排気弁４７を設けるので燃料に潤滑油を混合する必要が無く、排気に煤が発生する問題点が解消される効果があり、往復圧縮機２５は出力手段４のクランク軸４４による跳ねかけ潤滑を共用できる。
吸気弁４６は逆止弁としているが、弁駆動機構を設けて開閉することもできる。
水素の燃焼により水が発生し、少量の水または水蒸気は燃焼を助ける効果があるので排気性状が改善し、水は水蒸気になると約１７００倍に膨張して内燃機関の筒内圧力を増大するので出力増大効果がある。
燃料の遠心分離作用は、後述する実施例２にて、燃焼室を略球面状とし、吸気ポートをタンゼンシャルポート２３０とすることによる吸気の挙動と作用は、図２と図３にて、スワールの遠心分離作用により形成される各燃料濃度層の分布状況と点火時の火炎伝播は図５にて、前記遠心分離作用と燃焼性が向上する根拠となる燃料と空気（吸気）の特性は図６にて、前記掃気増幅手段５の作用、構成例は実施例４～６（図９～１７）にて説明する。 The operation of the internal combustion engine 1 of FIG. 1 is that the two-stroke internal combustion engine 1 consisting of a reciprocating compressor 25 performing a division cycle and an output unit 4 performs scavenging of scavenging air over the displacement of the internal combustion engine 1 by the reciprocating compressor 25; Scavenger inflow speed is increased by supplying swirl to the cylinder 41 from the sensial port 230 to generate swirl and increasing the scavenging amount supplied in the scavenging stroke, and the speed of swirl decreases due to the collision with the exhaust. The air flow is released as a short circuit scavenging air due to the blow through to make a strong swirl, and the strong swirl can be continued without the rotational movement of the swirl being inhibited by the shape of the combustion chamber until the end of the compression stroke.
Since the internal combustion engine 1 is a two-stroke internal combustion engine provided with an intake valve and an exhaust valve at the cylinder head, the scavenging efficiency is lower than that of the uniflow scavenging system with high scavenging efficiency. The combustion efficiency can be improved by improving the
The sum of the working angles of the vanes 83 of the displacement type hydraulic pressure supply means 8 and the working angles θc of the cam profile 881 of the cam 81 indicates the opening timing angle of the exhaust valve 47 by the hydraulic pressure. It becomes.
The check valve 875 of the hydraulic pressure auxiliary means 87 communicating with the hydraulic pressure hydraulic passage 88 for opening the exhaust valve 47 performs oil replenishment when hydraulic oil leaks, and thus has a lash adjuster function of the exhaust valve 47.
The valve drive mechanism of the exhaust valve 47 may be another drive mechanism as long as the valve arrangement is possible.
In the internal combustion engine 1, since the intake valve 46 and the exhaust valve 47 are provided in the cylinder head, there is no need to mix lubricating oil with fuel, and there is an effect that the problem of generating soot in the exhaust is eliminated. The splash lubrication by the crankshaft 44 of the output means 4 can be shared.
The intake valve 46 is a check valve, but can be opened and closed by providing a valve drive mechanism.
The combustion of hydrogen generates water, and a small amount of water or steam has the effect of assisting combustion, so the exhaust properties are improved, and when it is steam, it expands about 1700 times and increases the pressure in the cylinder of the internal combustion engine There is an output increase effect.
The centrifugal separation action of the fuel is as shown in FIG. 2 and FIG. 3 in the behavior and action of intake by making the combustion chamber substantially spherical and making the intake port the tangential port 230 in Example 2 described later. The distribution of the fuel concentration layers formed by the centrifugal separation action of the swirl and the flame propagation at the time of ignition in FIG. 5 are the characteristics of the fuel and air (intake air) that are the basis for the improvement of the centrifugal separation action and combustibility. The operation and configuration of the scavenging amplification means 5 will be described with reference to FIGS. 6 to 6 in FIGS.
図２は、実施例２（請求項１対応）の、略球面状の燃焼室にタンゼンシャルポートを設けた水素とガソリンを燃料とする内燃機関の平面図と周辺回路図である。
図２は、シリンダヘッドに吸気弁４６ｇと排気弁４７ｇを設けた２サイクル内燃機関１ｇにおいて、排気量より大きい容量の掃気を供給できる掃気供給手段を備え、
前記掃気供給手段は、前記内燃機関１ｇにて駆動する圧縮機２５ｇと、吸気通路である吸気流入通路２２ｇと吸気流出通路２３ｇの間に掃気増幅手段５ｇを設け、前記掃気増幅手段５ｇは、逆止弁５５ｇと前記逆止弁５５ｇの下流に設けた空気流量増幅器５０ｇから成り、前記空気流量増幅器５０ｇの駆動流通路５８ｇを前記圧縮機２５ｇの吐出口に連通し、更に、燃焼室を略球面状または略円錐状とし、前記燃焼室に放射状に吸気弁４６ｇと排気弁４７ｇを配置し、点火プラグ１１ｇを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポート２３０ｇとし、水素のように空気より密度が小さい燃料と火花点火式内燃機関の燃料であるガソリンを吸気系統および前記燃焼室に供給し、内燃機関１ｇの運転状況に応じて前記燃料の供給を制御する２サイクル内燃機関１ｇである。
流体供給手段７ｇの燃料タンク７５ｇに加圧貯蔵する水素を、吸気流出通路２３ｇに設けたインジェクタ１２ｇにて適時適量噴射し、掃気時に吹き抜けとならない吸気に予混合するために実施例４の図１３に示すタイミングに噴射する。
流体供給手段７ｇ２の燃料タンク７５ｇ２に貯蔵するガソリンを、燃焼室にインジェクタ１２ｇ２にて適時適量噴射する。
吸気弁４６ｇは逆止弁で、排気弁４７ｇは油圧通路８８ｇから供給される図示しない容積型油圧供給手段からの油圧にて弁シリンダ４７１ｇを作動して弁開閉を行う。
往復圧縮機２５ｇの構成と作用は実施例１と同じである。 FIG. 2 is a plan view and a peripheral circuit diagram of an internal combustion engine using a hydrogen and a gasoline as fuel, in which a tangential port is provided in a substantially spherical combustion chamber according to a second embodiment (corresponding to claim 1).
FIG. 2 shows scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the exhaust gas amount in a two-stroke internal combustion engine 1g provided with an intake valve 46g and an exhaust valve 47g in a cylinder head,
The scavenging air supply means includes a compressor 25g driven by the internal combustion engine 1g, and a scavenging air amplification means 5g between an intake air inflow passage 22g and an intake air outflow passage 23g, which are intake air passages. A stop valve 55g and an air flow amplifier 50g provided downstream of the check valve 55g communicate the drive flow passage 58g of the air flow amplifier 50g with the discharge port of the compressor 25g, and further the combustion chamber is substantially spherical. The intake valve 46g and the exhaust valve 47g are radially disposed in the combustion chamber, the spark plug 11g is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port is swirled in the cylinder. 230 g of tangential port to be used to supply fuel, which has a density smaller than that of air such as hydrogen, and gasoline, which is the fuel of a spark ignition type internal combustion engine, to the intake system and the combustion chamber A two-cycle internal combustion engine 1g for controlling the supply of the fuel in accordance with the operating condition of the internal combustion engine 1g.
The hydrogen to be pressurized and stored in the fuel tank 75g of the fluid supply means 7g is properly injected with an appropriate amount by the injector 12g provided in the intake and outlet passage 23g, and premixed with the intake air which does not blow when scavenging. Inject at the timing shown in.
The gasoline stored in the fuel tank 75g2 of the fluid supply means 7g2 is appropriately injected in an appropriate amount into the combustion chamber by the injector 12g2.
The intake valve 46g is a check valve, and the exhaust valve 47g operates the valve cylinder 471g with hydraulic pressure from a positive displacement hydraulic pressure supply means (not shown) supplied from the hydraulic pressure passage 88g to open and close the valve.
The configuration and operation of the reciprocating compressor 25g are the same as in the first embodiment.
図３は、前記実施例２（図２）の燃焼室の、圧縮行程の排気終了後の掃気中（Ｐ１）と圧縮行程の掃気終了後（Ｐ２）の各掃気挙動（スワール）の説明図である。
上図（Ｐ１）に示すように、排気行程が排気弁４７ｇの閉弁により終了し、２箇所のタンゼンシャルポート２３０ｇからシリンダ４１ｇに流入する吸気流は、シリンダ内で２本のスワールを発生し、図３では、２本のうちの一方のスワールを図示している。
下図（Ｐ２）に示すように、前記一方のスワールは、圧縮行程でピストン４２ｇにより軸方向に圧縮されてスワールは軸方向に圧縮変形し、略球面状の燃焼室ではスワール径の縮径が発生して、サイクロン効果のように強い遠心分離作用が発生する。
水素またはメタンのように空気より軽く、空気との密度差が大きい前記燃料は、前記強い遠心分離作用によりシリンダ軸対称に中心側に集まり、気体の密度勾配に応じて燃料は図５に示すように層状に分離して中心部に高濃度可燃層を形成し、燃焼室のシリンダ軸との交点近傍に設けた点火プラグまたはインジェクタの発火部に集まる。
ピストン４２ｇの頂面には、管状のバルブリセス４２１を設けることでスワールの乱れを抑制できる。
前記遠心分離効果は、内燃機関が高速になると遠心分離の作用時間が短くなるが、吸気流速が増大してスワールの回転が速くなるので、内燃機関の回転数による作用の影響は小さい。 FIG. 3 is an explanatory view of the scavenging behavior (swirl) during scavenging (P1) after exhaust of the compression stroke and after scavenging of the compression stroke (P2) in the combustion chamber of the second embodiment (FIG. 2). is there.
As shown in the above figure (P1), the exhaust stroke ends with the closing of the exhaust valve 47g, and the intake flow flowing into the cylinder 41g from the two tangential ports 230g generates two swirls in the cylinder. In FIG. 3, one of the two swirls is illustrated.
As shown in the lower drawing (P2), the one swirl is axially compressed by the piston 42g in the compression stroke, and the swirl is compressed and deformed in the axial direction, and a contraction of the swirl diameter occurs in the substantially spherical combustion chamber. As a result, strong centrifugal separation occurs like the cyclone effect.
The fuel which is lighter than air and has a large density difference with air such as hydrogen or methane gathers to the center side in the cylinder axis symmetry by the strong centrifugal separation action, and the fuel is shown in FIG. 5 according to the density gradient of the gas. The layers are separated into layers to form a high concentration combustible layer in the center and collected in the ignition part of the spark plug or injector provided near the intersection with the cylinder axis of the combustion chamber.
Disturbance of swirl can be suppressed by providing a tubular valve recess 421 on the top surface of the piston 42g.
The centrifugal effect shortens the action time of centrifugal separation when the internal combustion engine is at high speed, but since the intake flow velocity increases and the swirl rotation becomes fast, the influence of the action by the rotational speed of the internal combustion engine is small.
図４は、前記実施例２（図２）の燃焼室の排気終了時（Ｕ１）と圧縮終了時（Ｕ２）の容積図、と圧縮行程の試算による径方向容積占有率の変化の説明図（Ｕ３（Ｖ））である。図４の目的より、旋回流の図示および説明は省略する。
上図左（Ｕ１）は、排気終了時の径方向寸法を８等分した燃焼室の断面図であり、スワールにより左半分は手前方向に、右半分は奥方向にスワールが流れ上部の断面形状は略球面状の燃焼室で、周辺部と中央のシリンダ軸周辺との容積占有率の差はなだらかである。
上図右（Ｕ２）は、圧縮終了時の径方向に８等分した燃焼室の断面図であり、スワールが圧縮され、中央のシリンダ軸周辺と周辺部との容積占有率の差は拡大する。
下図（Ｕ３、Ｕ３Ｖ）は、縦軸が占有率、横軸がシリンダ径であり、縦軸の補助線は、シリンダ底面積を４等分する、シリンダ軸を中心とする同心円の各直径である。
下図左（Ｕ３）は、排気終了時と圧縮終了時の燃焼室の断面図から換算したそれぞれの占有率であり、排気終了時（破線）から圧縮終了時（実線）までの変化であり、圧縮行程で燃焼室形状の変化により径方向の前記等容積空間の占有率では、前記破線と実線で囲まれたハッチング部の高さに相当する変化量が発生し、シリンダ軸附近は前記破線の上部のハッチング分の占有率が増加し、周辺部は前記破線の下部のハッチング分の占有率が減少する。
尚、前記圧縮終了時（実線）は、二点鎖線で示す前記圧縮終了時の燃焼室の断面図（Ｕ２）を占有率に換算したものである。
下図右（Ｕ３Ｖ）は、シリンダ軸を中心として低面積が２５％毎増加する同心円により４等分割しているので、各領域の占有率の増減の試算値で、最外周の（７５～１００％）領域では、圧縮行程により占有率が半減し、一方シリンダ軸側の（０～２５％）領域の占有率は増加するので、各領域はシリンダ軸側にシフトし、スワールの旋回径が縮径するのでスワールの角速度が大きくなり高速回転するので、サイクロン効果のように遠心力の増大により強い遠心分離作用が発生し、点火時には高速の乱流火炎伝播が発生する。
下図右（Ｕ３Ｖ）の横軸から分かるように、シリンダの底面積は直径の二乗に比例するので、本願発明にて少量の燃料でもシリンダ軸側に燃料は集まり、点火プラグを設けたシリンダ軸付近に可燃層を確実に形成できる効果がある。
シリンダが垂直に設けられている場合は、空気より密度の小さい前記燃料は上方向に移動するので、更に点火プラグ附近に燃料が集まる。
以上のように、燃焼室形状は巨視的にみると圧縮行程にて略円筒状から略球面ドーム状に変化するので、吸気の径方向の燃料の層状分布は周辺部から中心部に移動し、スワールの旋回径が運動エネルギを保持して縮径するので角速度が大きくなり回転数が増大し、サイクロン効果のように遠心力の増大により強い遠心分離作用が発生する。 FIG. 4 is a volume diagram at the end of exhaust (U1) and at the end of compression (U2) of the combustion chamber of the second embodiment (FIG. 2), and an explanatory view of the change of the radial volume occupancy by trial calculation of the compression stroke ( U3 (V)). Illustration and description of the swirling flow are omitted for the purpose of FIG.
Top left (U1) is a cross-sectional view of the combustion chamber where the radial dimension at the end of exhaust is divided into eight equal parts, and the left half flows toward the front and the right half flows toward the back by swirl. Is a substantially spherical combustion chamber, and the difference in volume occupancy between the peripheral portion and the central cylinder axis is gentle.
Top right (U2) is a cross-sectional view of the combustion chamber divided into eight equal parts in the radial direction at the end of compression, and the swirl is compressed, and the difference in volume occupancy between the central cylinder axis and the peripheral part increases .
In the lower figure (U3, U3V), the vertical axis is the occupancy rate, the horizontal axis is the cylinder diameter, and the auxiliary line on the vertical axis is each diameter of concentric circles centered on the cylinder axis that divides the cylinder bottom area into four. .
The left figure (U3) in the figure below is the occupancy ratio converted from the cross-sectional view of the combustion chamber at the end of exhaust and at the end of compression, and is the change from the end of exhaust (dotted line) to the end of compression (solid line) With the change of the shape of the combustion chamber in the stroke, in the occupancy rate of the equal volume space in the radial direction, a variation corresponding to the height of the hatched portion surrounded by the broken line and solid line is generated. The occupancy rate of the hatching portion increases, and the occupancy rate of the hatching portion below the broken line decreases in the peripheral portion.
The end of the compression (solid line) is obtained by converting the cross-sectional view (U2) of the combustion chamber at the end of the compression indicated by a two-dot chain line into an occupancy ratio.
The right figure (U3V) in the figure below is divided into four equal areas by concentric circles where the low area increases by 25% around the cylinder axis, so the estimated value of the increase or decrease of the occupancy rate of each area In the region), the compression stroke reduces the occupancy by half, while the occupancy of the (0 to 25%) region on the cylinder axis increases, so each region shifts toward the cylinder axis, and the swirl diameter of the swirl decreases. Therefore, the angular velocity of the swirl increases and the motor rotates at a high speed. Therefore, as in the case of the cyclone effect, a strong centrifugal separation action occurs due to the increase of the centrifugal force, and high speed turbulent flame propagation occurs at the time of ignition.
As can be seen from the horizontal axis on the right (U3V) in the figure below, the bottom area of the cylinder is proportional to the square of the diameter, so even with a small amount of fuel according to the present invention There is an effect that the combustible layer can be formed with certainty.
In the case where the cylinder is vertically provided, the fuel having a density smaller than air moves upward, so that the fuel is further collected near the spark plug.
As described above, since the shape of the combustion chamber changes from a substantially cylindrical shape to a substantially spherical dome shape in a compression stroke as viewed macroscopically, the laminar distribution of fuel in the radial direction of intake moves from the peripheral portion to the central portion, Since the swirling diameter of the swirl holds a kinetic energy to reduce the diameter, the angular velocity increases and the number of rotations increases, and a strong centrifugal separation occurs due to the increase of the centrifugal force as in the cyclone effect.
図５は、前記実施例２（図１５）の、ＴＤＣの水素の各燃料濃度層の分布状況と点火時の高速火炎伝播の説明図である。
図５は、前記実施例２（図２）のＴＤＣのＸ－Ｘ断面図で、タンゼンシャルポート２３０ｇにより発生するスワールにて予混合気水素に遠心分離作用が働き、空気より密度が小さい水素（空気の約７％）燃料はシリンダ軸の中心方向に移動し、点火プラグ１１ｇを設けたシリンダ軸近傍に集まる。
図４で説明したように、角速度が大きくなるシリンダ軸近傍では、大きな遠心力が働くので空気より密度が小さい燃料が密度勾配により遠心分離して水素燃料の高濃度域（Ｆ１）を形成し、遠心分離途中の拡散した水素により、内側から順次燃料濃度の低い環状層（Ｆ２～Ｆ４）を形成する。
図３に示すように、圧縮行程で上図（Ｐ１）から下図（Ｐ２）に混合気を圧縮することにより、略球面状の燃焼室にスワール径が部分的に縮小するように圧縮されるので、図５の燃焼室頂点の点火プラグ１１ｇ付近に前記各層の境界面は収束する。 FIG. 5 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and high-speed flame propagation at the time of ignition in the second embodiment (FIG. 15).
FIG. 5 is a cross-sectional view of the TDC of Example 2 (FIG. 2) taken along the line XX, in which swirl generated by the tangential port 230 g causes centrifugal action on the premixed hydrogen, and hydrogen having a smaller density than air. The fuel moves toward the center of the cylinder axis and collects near the cylinder axis provided with the spark plug 11g.
As described in FIG. 4, in the vicinity of the cylinder axis where the angular velocity is large, a large centrifugal force works, and the fuel having a density smaller than that of air is centrifuged by the density gradient to form a high concentration region (F1) of hydrogen fuel, Annular layers (F2 to F4) with low fuel concentration are sequentially formed from the inside by the diffused hydrogen in the middle of the centrifugal separation.
As shown in FIG. 3, by compressing the mixture from the upper drawing (P1) to the lower drawing (P2) in the compression stroke, compression is performed so that the swirl diameter is partially reduced in the substantially spherical combustion chamber. The interface between the layers converges near the spark plug 11g at the top of the combustion chamber in FIG.
図５の内燃機関１ｇの燃焼行程点火時の層状分布の燃焼作用は、シリンダ軸付近の高濃度層（Ｆ１）に点火プラグ１１ｇにより点火して高濃度層（Ｆ１）内に火炎核を形成し、火炎伝播と燃焼による熱膨張と略同心円スワールにより高濃度燃料を拡散しながら、前記各層の燃料濃度の高い内側から濃度の低い外側の層に周方向に均一に乱流火炎伝播する。
前記高濃度層（Ｆ１）は、燃料の水素は空気との密度差が大きいので分離速度が大きく、高い濃度の高濃度層を形成し、図６に示すように燃焼範囲（約４～７５％）が大きいので、確実な点火ができる。
圧縮行程でのスワールにより、遠心分離の初期分離状態であるので、予混合による燃料と空気（酸素）との混合拡散は、高濃度層（Ｆ１）、中濃度層（Ｆ２）、低濃度層（Ｆ３）、超低濃度層（Ｆ４）の燃料混合層であり、燃焼の火炎伝播は超低濃度層（Ｆ４）の外周では低温燃焼または壁面附近で消炎するのでノッキング現象が抑制され、前記超低濃度層Ｆ４での消炎により未燃焼燃料が減少するので、内燃機関の冷却損失の抑制による燃焼効率の向上の効果がある。
従って、成層燃焼により従来技術より少量の燃料による燃焼効率のよいリーンバーンエンジンとなる。
尚、以降の実施例の遠心分離により形成される各層の符号は、理解が容易なように実施例２（図５）と同じ層符号（Ｆ１～Ｆ４）を用いて説明する。
ガソリン供給時は供給タイミングにより、中濃度層（Ｆ２）または低濃度層（Ｆ３）に拡散するので、前記水素の高速火炎伝播により燃焼が促進され、水素の高速燃焼により燃焼室の圧力が上昇しＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）により予混合ガソリンの燃焼性を改善する効果がある。
実施例２がディーゼル機関の場合は、燃料の圧縮着火が水素の高濃度で始まるので、燃焼が促進されるので回転数の増大により出力増大効果がある。 The combustion action of the stratified distribution at the time of combustion stroke ignition of the internal combustion engine 1g of FIG. 5 forms a flame kernel in the high concentration layer (F1) by igniting the high concentration layer (F1) near the cylinder axis by the spark plug 11g. While diffusing high concentration fuel by thermal expansion due to flame propagation and combustion and approximately concentric swirl, the turbulent flame is uniformly propagated circumferentially from the inner side with high fuel concentration of each layer to the outer layer with low concentration.
The high concentration layer (F1) has a large separation speed because hydrogen of the fuel has a large density difference with air, and forms a high concentration layer of high concentration, and the combustion range (about 4 to 75% as shown in FIG. 6) ) Is large, you can do a reliable ignition.
Because of the initial separation state of centrifugation due to the swirl in the compression stroke, the mixed diffusion of fuel and air (oxygen) by premixing is high concentration layer (F1), middle concentration layer (F2), low concentration layer ( F3) Fuel mixing layer of ultra-low concentration layer (F4), flame propagation of combustion is low temperature combustion around the outer periphery of ultra-low concentration layer (F4) or flame extinction near wall surface, so knocking phenomenon is suppressed, Since the unburned fuel is reduced due to the extinction in the concentration layer F4, there is an effect of improving the combustion efficiency by suppressing the cooling loss of the internal combustion engine.
Thus, stratified charge combustion results in a lean burn engine with good combustion efficiency with less fuel than the prior art.
The reference numerals of the layers formed by centrifugation in the following examples are described using the same layer reference (F1 to F4) as that of the second embodiment (FIG. 5) for easy understanding.
At the time of gasoline supply, it diffuses to the medium concentration layer (F2) or the low concentration layer (F3) depending on the supply timing, so the high-speed flame propagation of hydrogen promotes combustion, and the high-speed combustion of hydrogen raises the pressure in the combustion chamber The SPCCI engine (compression ignition combustion with spark control) has the effect of improving the flammability of premixed gasoline.
In the case of the diesel engine according to the second embodiment, since the compression ignition of the fuel starts with a high concentration of hydrogen, the combustion is promoted, and the power increase effect is obtained by the increase of the rotational speed.
図６は、前記実施例２（図２）の内燃機関等の燃料と空気の特性図で、燃料は燃焼範囲と密度、空気は組成割合と密度を示す。
図の縦軸は気体の密度（Ｋｇ／ｍ３）、横軸は、燃料は燃焼範囲、空気は組成割合（Ｖｏｌ％）である。
空気の主な組成割合は約２１％の酸素と約７８％の窒素であり、酸素の密度は空気の約１．０７倍、窒素の密度は空気の約０．９３倍で、空気との密度差が小さく空気中に拡散した酸素と窒素は、本発明の短時間の遠心分離では分離困難である。
燃料である水素の密度は空気の約０．０７倍であり短時間の遠心分離で分離を開始し、メタンの密度は空気の約０．５３倍であり水素ほど容易に遠心分離しないが、これらの燃料は遠心分離で密度勾配により形成される各層の最内側に分離される。
プロパンの密度は空気の約１．４７倍であり、ガソリンの密度は空気の約２．７１倍であるので、これらの燃料は遠心分離で密度勾配により形成される前記最内層より外側の濃度層に分離される。
燃料の横軸の燃焼範囲は、ＥＧＲガスに少量含まれる一酸化炭素を除けば、水素以外の燃料は小さい燃料範囲で燃焼するが、水素は４．１～７５（Ｖｏｌ％）と大きい範囲で燃焼する。
背景技術で述べたように、最小着火エネルギ（ｍｊ）が水素（０．０２）はガソリン（０．２４）より小さく、最大燃焼速度（ｃｍ／ｓ）が水素（３４６）はガソリン（４２）より大きいので、水素は着火性がよく、爆風圧が大きい利点があるので燃焼の起爆剤に適しているが、燃料としては発熱量が小さいのでエネルギ密度が小さい問題点がある。 FIG. 6 is a characteristic diagram of fuel and air of the internal combustion engine etc. of the second embodiment (FIG. 2), wherein the fuel shows the combustion range and the density, and the air shows the composition ratio and the density.
The vertical axis of the figure is the density of gas (Kg / m 3), the horizontal axis is the combustion range of the fuel, and the composition ratio (Vol%) of the air.
The main composition ratio of air is about 21% oxygen and about 78% nitrogen, the density of oxygen is about 1.07 times that of air, the density of nitrogen is about 0.93 times that of air, and the density with air The differences between oxygen and nitrogen, which are small and diffused in the air, are difficult to separate in the short time centrifugation of the present invention.
The density of hydrogen, which is a fuel, is about 0.07 times that of air and separation is started by a short centrifugation, and the density of methane is about 0.53 times that of air, which is not as easily centrifuged as hydrogen. Fuel is separated in the innermost of the layers formed by density gradient by centrifugation.
Since the density of propane is about 1.47 times that of air and the density of gasoline is about 2.71 times that of air, these fuels have a concentration layer outside of the innermost layer formed by density gradient in centrifugation. Separated into
With the exception of carbon monoxide, which is contained in a small amount in the EGR gas, the combustion range of the fuel on the horizontal axis of the fuel burns in a small fuel range except for hydrogen, but hydrogen ranges as large as 4.1 to 75 (Vol%) To burn.
As mentioned in the background art, the minimum ignition energy (mj) is less than hydrogen (0.02) than gasoline (0.24) and the maximum burn rate (cm / s) is less than hydrogen (346) than gasoline (42) Because it is large, hydrogen is suitable for combustion initiators because it has good ignition performance and high blast pressure. However, it has a problem that energy density is small because its calorific value is small as fuel.
前記図４（Ｕ３Ｖ）に示すように、燃焼室の中心（シリンダ軸附近）にシリンダ径の５０％の燃焼能域を形成するには水素は２５％（Ｖｏｌ％）必要となる。
点火プラグによる点火に必要な可燃層を仮にシリンダ径の２０％の直径と仮定すると、必要な可燃層はシリンダ容積の４％となり、前記可燃層の水素濃度を水素の燃焼範囲の下限の４．１％とすると、必要な水素は、前記可燃層のシリンダ容積の４％と、前記水素の燃焼範囲の下限の４．１％の積である約０．１６％となり、見かけの濃度は極めて低い濃度であっても、試算上は局部的に水素の可燃層が形成できる。
前記試算は周辺に拡散する水素を無視した高濃度層（Ｆ１）のみの試算であり、遠心分離作用の初期段階であるので水素は各層に拡散して存在するので、実際に必要な水素は前記試算値より大きくなる。
以上より、水素より分子量が大きい主燃料を少量の水素で局部的に可燃層を形成し、確実に点火して高速燃焼できる成層燃焼の２サイクル内燃機関（実施例１～１３）と４サイクル内燃機関（実施例１４，１５）ができる。 As shown in FIG. 4 (U3V), 25% (Vol%) of hydrogen is required to form a combustion area of 50% of the cylinder diameter at the center of the combustion chamber (around the cylinder axis).
Assuming that the combustible layer necessary for ignition by the spark plug is 20% of the cylinder diameter, the necessary combustible layer is 4% of the cylinder volume, and the hydrogen concentration of the combustible layer is the lower limit of the hydrogen combustion range of 4. Assuming 1%, the required hydrogen is about 0.16% which is the product of 4% of the cylinder volume of the combustible layer and 4.1% of the lower limit of the hydrogen combustion range, and the apparent concentration is extremely low. Even in the case of concentration, it is possible to form a combustible layer of hydrogen locally in the estimation.
The above-mentioned trial calculation is a trial calculation of only the high concentration layer (F1) neglecting hydrogen diffused to the periphery, and since hydrogen is diffused and present in each layer because it is the initial stage of the centrifugal separation action, actually necessary hydrogen is said It becomes larger than the estimate.
From the above, a two-stroke internal combustion engine with stratified combustion (Examples 1 to 13) and a four-stroke internal combustion that can locally burn a main fuel with a molecular weight larger than hydrogen and form a flammable layer locally with a small amount of hydrogen. An organization (Examples 14 and 15) can be done.
図７は、実施例３（請求項１対応）の往復圧縮機、吸排気の弁と通路の配置を示す３気筒内燃機関の平面図と、掃気増幅手段と容積型油圧供給手段等の周辺回路図である。
図７は、シリンダヘッドに吸気弁４６ｊと排気弁４７ｊを設けた２サイクル内燃機関１ｊにおいて、排気量より大きい容量の掃気を供給できる掃気供給手段を備え、前記掃気供給手段は、前記内燃機関１ｊにて駆動する圧縮機２５ｊと、吸気通路である吸気流入通路２２ｊと吸気流出通路２３ｊの間に掃気増幅手段５ｊを設け、前記掃気増幅手段５ｊは、逆止弁５５ｊと前記逆止弁５５ｊの下流に設けた空気流量増幅器５０ｊから成り、前記空気流量増幅器５０ｊの駆動流通路５８ｊを前記圧縮機２５ｊの吐出口に連通し、更に、燃焼室を略球面状または略円錐状とし、前記燃焼室に放射状に吸気弁４６ｊと排気弁４７ｊを配置し、点火プラグ１１ｊを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポート２３０ｊとし、水素、メタンのように空気より密度が小さい燃料と火花点火式内燃機関の燃料を吸気系統および前記燃焼室に供給し、内燃機関１ｊの運転状況に応じて前記燃料の供給を制御する３気筒２サイクル内燃機関１ｊである。
各気筒の掃気供給手段の構成は実施例１と同じであり、往復圧縮機２５ｊと、掃気増幅手段５ｊを吸気流入通路２２ｊと吸気流出通路２３ｊの間に気筒毎に設けている。
吸気流出通路２３ｊを３気筒分連通して一体のマニフォルドとすることもできるが、掃気と過給の性能が低下する。
容積型油圧供給手段８ｊは、前記実施例１の容積型油圧供給手段８に等間隔にベーン８３ｊを増設することによりカム８１１を共用して３回路の油圧通路８８（ｊ１～ｊ３）を配置できるので、簡素な構造で、油圧手段の信頼性が高く、小型で安価に製作できる。 FIG. 7 is a plan view of a three-cylinder internal combustion engine showing the arrangement of the reciprocating compressor, intake and exhaust valves and passages according to the third embodiment (corresponding to claim 1), and peripheral circuits such as scavenging air amplification means and positive displacement hydraulic pressure supply means. FIG.
FIG. 7 shows scavenging gas supply means capable of supplying scavenging gas having a capacity larger than the exhaust gas amount in a two-stroke internal combustion engine 1j provided with an intake valve 46j and an exhaust valve 47j in a cylinder head. The scavenging amplification means 5j is provided between the compressor 25j driven at the same time and the intake inflow path 22j and the intake and outflow path 23j, which are intake paths, and the scavenging amplification means 5j includes a check valve 55j and the check valve 55j. The air flow amplifier 50j is provided downstream, the drive flow passage 58j of the air flow amplifier 50j is in communication with the discharge port of the compressor 25j, and the combustion chamber has a substantially spherical or substantially conical shape. The intake valve 46 j and the exhaust valve 47 j are radially disposed in the cylinder, the spark plug 11 j is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber, and the intake port generates swirl in the cylinder The fuel is supplied to the intake system and the combustion chamber, such as hydrogen and methane, which have a smaller density than air and fuel of a spark ignition type internal combustion engine, according to the operating condition of the internal combustion engine 1j. A three-cylinder two-stroke internal combustion engine 1j that controls the supply of
The configuration of the scavenging gas supply means of each cylinder is the same as that of the first embodiment, and the reciprocating compressor 25j and the scavenging gas amplification means 5j are provided for each cylinder between the intake inflow passage 22j and the intake outflow passage 23j.
The intake and outflow passages 23j may be communicated by three cylinders to form an integral manifold, but the performance of scavenging and supercharging may be reduced.
The positive displacement hydraulic pressure supply means 8j can arrange the three hydraulic pressure passages 88 (j1 to j3) by sharing the cam 811 by adding the vanes 83j to the positive displacement hydraulic pressure supply means 8 of the first embodiment at equal intervals. Therefore, with a simple structure, the hydraulic means can be manufactured highly reliable, compact and inexpensive.
図８は、実施例３（図７）のＪ－Ｊ断面の吸気弁と排気弁の冷却手段を設けた前記内燃機関１ｊの断面図である。
図８は、前記図７のＪ－Ｊ断面のガス交換弁である吸気弁４６ｊと排気弁４７ｊの断面図で、前記吸気弁４６ｊはリフト逆止弁で掃気の圧力により開弁し、排気弁４７ｊの弁駆動機構は弁シリンダ４７１ｊに油圧通路８８ｊ３から供給される油圧により開弁し、前記吸気弁４６ｊはリフト逆止弁に送風ブレード４６６を、排気弁４７ｊは油圧ピストンに一体の送風ブレード４７６を設け、送風ブレードのポンプ作用により連通管４６４から吸気の一部を前記吸気の逆止弁と排気弁の弁シリンダ４７１ｊの当接部空間に送り冷却を行う。弁シリンダ４７１ｊのピストンに設けた送風ブレード４７６の往復動と各逆止弁（４７３、４７７、４７８）の作用により、前記送風ブレード４６６の上部と下部の空間をポンプ室とするポンプ作用により、冷却用の吸気を送り込み、逆止弁４６３、４６８を通って止弁４６７から大気に解放する。
排気の弁シリンダは、連通管４６４を通った吸気の一部が吸気導管４４５を通って前記冷却を行う。 FIG. 8 is a cross-sectional view of the internal combustion engine 1j provided with cooling means for the intake valve and the exhaust valve of the JJ cross section of the third embodiment (FIG. 7).
FIG. 8 is a cross-sectional view of the intake valve 46j and the exhaust valve 47j which are gas exchange valves of the JJ cross section of FIG. 7, and the intake valve 46j is opened by the pressure of scavenging with a lift check valve. The valve drive mechanism 47j is opened by the hydraulic pressure supplied from the hydraulic pressure passage 88j3 to the valve cylinder 471j, and the intake valve 46j is a lift check valve, and the exhaust valve 47j is a blow blade 476 integrated with the hydraulic piston. A part of the intake air is sent from the communication pipe 464 to the contact space of the check valve of the intake valve and the valve cylinder 471j of the exhaust valve by the pump action of the blower blade to perform cooling. The reciprocating action of the blower blade 476 provided on the piston of the valve cylinder 471j and the action of the respective check valves (473, 477, 478), the pump action which makes the space above and below the blower blade 466 a pump chamber Supply air and release it from the stop valve 467 to the atmosphere through the check valves 463, 468.
In the exhaust valve cylinder, a portion of the intake air passing through the communication pipe 464 performs the cooling through the intake conduit 445.
図９は、実施例４（請求項１対応）の逆止弁と空気流量増幅器から成る掃気増幅手段と、クランク軸より位相が進んだ往復圧縮機とを備えた２サイクル内燃機関の構成概念の説明図である。
図９は、掃気供給手段として、内燃機関１ｄにて駆動する圧縮機である往復圧縮機２５ｄと、吸気通路である吸気流入通路２２ｄと吸気流出通路２３ｄの間に掃気増幅手段５を設け、前記掃気増幅手段５は、逆止弁５５と前記逆止弁の下流に設けた空気流量増幅器５０から成り、前記空気流量増幅器５０の駆動流通路５８を前記往復圧縮機２５ｄの吐出弁２５７ｄの吐出口に連通する請求項１に記載の２サイクル内燃機関１ｄである。
出力手段４ｄより狭角のθｄだけ位相が進んだ往復圧縮機２５ｄは、吸入弁２５６ｄと吐出弁２５７ｄに逆止弁を備え、連結棒２５３ｄを出力手段４ｄの連結棒４３ｄにピギー接続してストロークを出力手段４より短くすることにより高速回転に対応ができ、シリンダ２５１ｄの直径は出力手段４ｄのシリンダ４１ｄの直径より小さくし、掃気の吹き抜けによる充填効率の低下や排気量以上の吸気が必要な過給は前記空気流量増幅器５０の流量増幅作用により、小さい容量の前記往復圧縮機２５ｄで対応できる。
出力手段４ｄの吸入弁４６ｄは逆止弁とし、クランク軸４４ｄに設けた駆動車４０１が伝動媒体４０３を介して同じ有効径（φＤ）の従動車４０２を回転し、前記従動車４０２に設けたカム４０８にて排気弁４７を開閉する。
流体供給手段７の燃料タンク７５ｄに貯蔵する重油または軽油をサプライポンプ１３１で加圧し、コモンレール１４１を通ってインジェクタ１２ｄにて適時燃焼室に供給する。
前記往復圧縮機２５ｄの吸入弁２５６ｄと掃気増幅手段５の上流はエアクリーナ２１ｄに連通し、出力手段４ｄの排気弁４７ｄは排気通路３１ｄを介して消音器３３ｄの上流の排気浄化装置３２ｄに連通する。 FIG. 9 shows the construction of a two-stroke internal combustion engine provided with scavenging amplification means comprising a check valve and an air flow amplifier according to a fourth embodiment (corresponding to claim 1), and a reciprocating compressor whose phase is advanced from the crankshaft. FIG.
FIG. 9 shows, as scavenging air supply means, a reciprocating compressor 25d, which is a compressor driven by the internal combustion engine 1d, and a scavenging air amplification means 5 between the intake inflow passage 22d and the intake outflow passage 23d, which are intake passages; The scavenging air amplification means 5 comprises a check valve 55 and an air flow amplifier 50 provided downstream of the check valve, and the drive flow passage 58 of the air flow amplifier 50 is a discharge port of the discharge valve 257d of the reciprocating compressor 25d. It is a two-stroke internal combustion engine 1d according to claim 1, which communicates with the engine.
The reciprocating compressor 25d whose phase is advanced by θd at a narrower angle than the output means 4d has a check valve in the suction valve 256d and the discharge valve 257d, and carries out a stroke by connecting the connecting rod 253d to the connecting rod 43d of the output means 4d. The length of the cylinder 251d can be made smaller than the diameter of the cylinder 41d of the output means 4d, and the intake efficiency must be reduced by the scavenging air flow and the intake efficiency must be greater than the exhaust capacity. Supercharging can be handled by the reciprocating compressor 25d with a small capacity by the flow amplification function of the air flow amplifier 50.
The suction valve 46 d of the output means 4 d is a check valve, and the drive wheel 401 provided on the crankshaft 44 d rotates the driven wheel 402 of the same effective diameter (φD) via the transmission medium 403 and provided on the driven wheel 402 The exhaust valve 47 is opened and closed by the cam 408.
The heavy oil or light oil stored in the fuel tank 75 d of the fluid supply means 7 is pressurized by the supply pump 131 and supplied to the combustion chamber through the common rail 141 by the injector 12 d as appropriate.
The suction valve 256d of the reciprocating compressor 25d and the upstream of the scavenging air amplification means 5 communicate with the air cleaner 21d, and the exhaust valve 47d of the output means 4d communicates with the exhaust purification device 32d upstream of the silencer 33d via the exhaust passage 31d. .
図９の内燃機関１ｄの作用は、往復圧縮機２５で発生する圧縮空気を掃気増幅手段５の空気流量増幅器５０に駆動流として供給し、前記空気流量増幅器５０にて吸気を流量増幅して大気圧以上の圧力で吸気を出力手段４ｄの吸気弁４６ｄに供給して内燃機関１ｄの掃気を行う。
空気流量増幅器５０の下流側の掃気圧力が高くなりすぎると、エアクリーナ２１ｄに吸気が逆流し、更に駆動流により逆流方向に流量増幅する逆流量増幅現象が発生するのを逆止弁５５で防止する。このように空気流量増幅器５０の下流側の掃気圧力が高くなりすぎると、前記逆止弁が作動し、駆動流が直接掃気に流入するので、掃気が高圧となり過給作用が発生する。
前記空気流量増幅器５０の流量増幅比に応じて吸気を流量増幅するので、往復圧縮機２５ｄの吐出量は内燃機関１ｄの排気量より小さい容量でよい。
出力手段４ｄのシリンダ４１ｄより往復圧縮機２５ｄのシリンダ２５１ｄの直径が小さく、ストロークも短い小型で安価な往復圧縮機２５ｄで十分に掃気ができ、往復圧縮機２５ｄの潤滑を出力手段４ｄの跳ね掛け潤滑を共用できるので信頼性が高い。
簡素な構成の前記掃気増幅手段５で吹き抜けを補填して余りある掃気を供給することにより、完全なガス交換と駆動流による掃気の加圧により過給ができる。
半径ＳＲｄの略球面状燃焼室に、タンゼンシャルポート２３０ｄによるスワールの排気と衝突して掃気を行う初期流入吸気を、吹き抜けとして排気と流出し、強いスワールによる遠心分離により水素等の低密度気体燃料を発火部に集中して燃焼性を向上できる効果がある。
排気弁４７ｄの弁駆動は前記実施例３と同じであるので説明を省略する。
掃気増幅手段５の空気流量増幅器５０と逆止弁５５の構成例の構成説明を図１０にて、掃気増幅手段５、往復圧縮機、および出力手段の作用の動作説明を図１１にて説明する。
内燃機関１の、タイミングチャートを図１２にて、タイミングチャートと試算による筒内圧力を図１３にて、高速回転時の試算によるＰＶ線図を図１４にて説明する。
本実施例４の内燃機関１はディーゼル機関であるが、火花点火式内燃機関でもよい。 The operation of the internal combustion engine 1d shown in FIG. 9 supplies compressed air generated by the reciprocating compressor 25 as a drive flow to the air flow amplifier 50 of the scavenging amplifier 5, and the air flow amplifier 50 amplifies the flow of intake air to increase The intake air is supplied to the intake valve 46d of the output means 4d at a pressure higher than the atmospheric pressure to scavenge the internal combustion engine 1d.
If the scavenging pressure on the downstream side of the air flow amplifier 50 becomes too high, the check valve 55 prevents the occurrence of a reverse flow amplification phenomenon in which the intake air flows back to the air cleaner 21d and the flow is amplified in the reverse direction by the drive flow. . As described above, when the scavenging pressure on the downstream side of the air flow amplifier 50 becomes too high, the check valve is activated and the drive flow directly flows into the scavenging air, so that the scavenging air has a high pressure and a supercharging action occurs.
Since the flow rate of the intake air is amplified according to the flow rate amplification ratio of the air flow rate amplifier 50, the displacement of the reciprocating compressor 25d may be smaller than the displacement of the internal combustion engine 1d.
The diameter of the cylinder 251d of the reciprocating compressor 25d is smaller than the cylinder 41d of the output means 4d, and the stroke can be sufficiently scavenged with a small and inexpensive reciprocating compressor 25d having a short stroke, and the lubrication of the reciprocating compressor 25d is splashed by the output means 4d. High reliability because lubrication can be shared.
Supercharging can be performed by complete gas exchange and pressurization of the scavenging air by the driving flow by supplying the scavenging air having a surplus by compensating the blow-by with the scavenging air amplification means 5 having a simple configuration.
Initial inflowing intake air which scavenges by collision with exhaust of swirl by tangential port 230d into a substantially spherical combustion chamber of radius SRd as exhaust and outflow as blowout, and low density gas such as hydrogen by centrifugal separation by strong swirl There is an effect that fuel can be concentrated on the igniter to improve the combustibility.
Since the valve drive of the exhaust valve 47d is the same as that of the third embodiment, the description will be omitted.
A configuration of the air flow amplifier 50 and the check valve 55 of the scavenging amplification means 5 will be described with reference to FIG. 10, and the operation of the scavenging amplification means 5, the reciprocating compressor, and the output means will be described with reference to FIG. .
The timing chart of the internal combustion engine 1 will be described with reference to FIG. 12, the in-cylinder pressure based on the timing chart and trial calculation will be described with reference to FIG. 13, and a PV diagram based on trial calculation at high speed rotation will be described with reference to FIG.
Although the internal combustion engine 1 of the fourth embodiment is a diesel engine, it may be a spark ignition internal combustion engine.
図１０は前記実施例４（図９）の掃気増幅手段の構成例で、空気流量増幅器の流量増幅比の小さい順にエジェクタ（Ａ）、従来技術（特開２０１６－１２５４２１）のフロートランスベクタ（Ｂ）とトランスベクタ（Ｃ）と逆止弁の構成説明図である。
逆止弁は、リードバルブ５５１、リフト逆止弁５５５（Ｃ）または他の逆止弁でもよく、応答性、耐圧等より内燃機関の仕様により選択できる。
空気流量増幅器は、主に流量増幅比により選定し、内燃機関の運転状況が変動する場合は高速領域等で掃気増幅手段５での圧力損失が増大して運転効率が低下するので、前記（Ｂ）、（Ｃ）に示す前記従来技術（特開２０１６－１２５４２１）の運転領域が大きいノズル開口面積可変型が好ましい。 FIG. 10 shows a configuration example of the scavenging air amplification means of the fourth embodiment (FIG. 9), in which the flow transvectors (B) of the ejector (A) and the prior art (Japanese Patent Laid-Open No. 2016-125421) are arranged in ascending order of flow amplification ratio of the air flow amplifier. It is structural explanatory drawing of a trans | transformer vector (C) and a non-return valve.
The check valve may be a reed valve 551, a lift check valve 555 (C) or another check valve, which can be selected according to the specification of the internal combustion engine from the response, the pressure resistance, and the like.
The air flow amplifier is mainly selected by the flow amplification ratio, and when the operating condition of the internal combustion engine fluctuates, the pressure loss in the scavenging amplification means 5 increases in the high speed region etc. and the operation efficiency decreases. The nozzle opening area variable type having a large operation area of the prior art (Japanese Patent Application Laid-Open No. 2016-125421) shown in the above (c) is preferable.
図１１は、前記実施例４（図９）の内燃機関１ｄの、排気行程初期（Ｓ１）、掃気行程（Ｓ２）、および圧縮行程（Ｓ３）の掃気増幅手段５の動作説明図である。
排気行程初期（Ｓ１）は、燃焼が終了した排気を排気弁４７ｄの開弁により排気を開始し、往復圧縮機２５は圧縮の初期であるので掃気増幅手段５に駆動流通路５８から圧縮空気を供給し流量増幅を開始するが、排気の圧力が高い場合は逆止弁である吸気弁４６ｄは開弁しない。
掃気行程（Ｓ２）は、排気が進行し排気圧が低下し、往復圧縮機２５の駆動流圧力が上昇して掃気増幅手段５による流量増幅により吸気流出通路２３ｄの掃気圧力が上昇して前記吸気弁４６ｄは開弁すると、前記流量増幅は更に進行し、シリンダ４１ｄの掃気を開始する。
圧縮行程（Ｓ３）は、排気弁４７ｄが閉弁し、吸気弁４６ｄから掃気が供給されるのでシリンダ４１ｄが大気圧より高くなり、吸気流出通路２３ｄの圧力が上昇すると逆流流量増幅現象により逆止弁５５が閉弁し、往復圧縮機２５の圧縮空気は直接吸気流出通路２３ｄに供給され吸気弁４６ｄより供給されるので過給効果が発生する。 FIG. 11 is an operation explanatory view of the scavenging amplification means 5 of the initial exhaust stroke (S1), scavenging stroke (S2), and compression stroke (S3) of the internal combustion engine 1d of the fourth embodiment (FIG. 9).
At the beginning of the exhaust stroke (S1), the exhaust after the combustion is started to be exhausted by opening the exhaust valve 47d, and the reciprocating compressor 25 is at the initial stage of compression. The supply flow rate amplification is started, but when the pressure of the exhaust is high, the intake valve 46d which is a check valve does not open.
In the scavenging stroke (S2), the exhaust proceeds and the exhaust pressure decreases, and the driving flow pressure of the reciprocating compressor 25 rises, and the scavenging pressure of the intake outflow passage 23d rises by the flow amplification by the scavenging amplification means 5 and the intake When the valve 46d is opened, the flow rate amplification further proceeds to start scavenging of the cylinder 41d.
In the compression stroke (S3), since the exhaust valve 47d is closed and scavenging air is supplied from the intake valve 46d, the cylinder 41d becomes higher than the atmospheric pressure and the pressure in the intake and outlet passage 23d increases. The valve 55 is closed, and the compressed air of the reciprocating compressor 25 is directly supplied to the intake / outlet passage 23d and supplied from the intake valve 46d, so that a supercharging effect occurs.
図１２は、前記実施例４（図９）の内燃機関のタイミングダイアグラムである。
図１２の燃焼行程（Ｂ）は、ＴＤＣでの燃焼開始から始まりＢＤＣよりΔＥだけ位相が手前の排気弁４７ｄの開弁で終了し、排気行程（Ｅ）は前記排気弁４７ｄの開弁から始まり排気弁４７ｄの閉弁により終了し、圧縮行程は前記排気弁４７ｄの閉弁から始まり前記ＴＤＣで終了する。掃気行程は、ＢＤＣよりΔＳだけ遅れて前記排気行程の中盤から始まり、前記排気弁４７ｄの閉弁よりＣｓだけ位相が遅れて終了し、前記ΔｓとＤｓは排気と掃気の圧力差により作動する逆止弁である吸気弁４６ｄの弁の開閉により決まるタイミング値であり内燃機関１ｄの運転状況等により変動する。
図１２から分かるように排気弁４７ｄの作動により排気行程は、シリンダストロークに対して対象にする必要が無いので燃焼行程は十分な膨張仕事ができ、排気行程と圧縮行程に重複する掃気行程は、排気が十分排出されて略大気圧に減圧されたシリンダ４１ｄに掃気を流入して効率よくガス交換ができ、排気弁４７ｄの閉弁後に掃気を更に供給することにより過給効果が発生する。 FIG. 12 is a timing diagram of the internal combustion engine of the fourth embodiment (FIG. 9).
The combustion stroke (B) in FIG. 12 starts from the start of combustion at TDC and ends by opening the exhaust valve 47d whose phase is ΔE earlier than BDC, and the exhaust stroke (E) starts from opening the exhaust valve 47d. The compression stroke ends with the closing of the exhaust valve 47d, and the compression stroke starts from the closing of the exhaust valve 47d and ends with the TDC. The scavenging stroke starts from the middle of the exhaust stroke later by ΔS than BDC and ends in phase delayed by Cs from closing of the exhaust valve 47d, and the Δs and Ds are reversely operated by the pressure difference between the exhaust and scavenging air. It is a timing value determined by the opening and closing of the valve of the intake valve 46d which is a stop valve, and fluctuates depending on the operating condition of the internal combustion engine 1d and the like.
As can be seen from FIG. 12, the exhaust stroke does not have to be targeted with respect to the cylinder stroke by the operation of the exhaust valve 47d, so the combustion stroke can perform sufficient expansion work, and the scavenging stroke overlapping the exhaust stroke and the compression stroke is Exhaust gas is sufficiently discharged and scavenging air flows into the cylinder 41d reduced to substantially atmospheric pressure to perform efficient gas exchange, and by further supplying scavenging air after the exhaust valve 47d is closed, a supercharging effect is generated.
図１３は、実施例４（図９）の往復圧縮機と掃気増幅手段を設けた内燃機関の各部のタイミングチャートと試算による筒内圧力である。
図１３の横軸は、２サイクル内燃機関（３６０°）のクランク角変位量であり、縦軸の各項目は、上段より作動行程は前記図１２の内燃機関１ｄのタイムチャート（帯グラフ）、次は燃料供給のタイミングチャート（実施例４と異なる（火花点火式内燃機関）と（掃気予混合）を含む）、次は出力手段４ｄと往復圧縮機２５ｄの各要素のタイミングチャートで、往復圧縮機２５ｄのピストン２５２ｄの変異のハッチング部は往復圧縮機２５２ｄの吸入量を示す。
最下段は以上の結果として内燃機関１ｄの出力を発生するシリンダ４１ｄの筒内圧力の変動の試算値である。
燃料供給は、各内燃機関をハッチングのタイミングに燃料供給することにより、掃気の吹き抜けによる未燃焼燃料の流出を防止できる。
最下段のシリンダ４１ｄの筒内圧力の変動は、ディーゼル機関である内燃機関１ｄの高速高負荷時の燃焼（複合サイクル）であり、火花点火式内燃機関であっても図１３と燃焼方法以外は同じであり、タイミングチャートは掃気タイミングが過給圧等により多少変動するが大差はない。 FIG. 13 shows in-cylinder pressures estimated by calculation and timing charts of respective parts of an internal combustion engine provided with the reciprocating compressor and the scavenging air amplification means of the fourth embodiment (FIG. 9).
The horizontal axis of FIG. 13 is the crank angle displacement amount of the two-stroke internal combustion engine (360 °), and each item of the vertical axis is the time chart (band graph) of the internal combustion engine 1d of FIG. Next is a timing chart of fuel supply (different from the fourth embodiment (including spark ignition type internal combustion engine) and (including scavenging air mixing)), next is a timing chart of each element of the output means 4d and the reciprocating compressor 25d. The hatched portion of the variation of the piston 252d of the machine 25d indicates the suction amount of the reciprocating compressor 252d.
The lowermost portion is a trial calculation value of the in-cylinder pressure fluctuation of the cylinder 41d that generates the output of the internal combustion engine 1d as a result of the above.
The fuel supply can prevent the outflow of unburned fuel due to scavenging air by fueling the internal combustion engines at the timing of hatching.
The fluctuation of the pressure in the lowermost cylinder 41d is the combustion (combined cycle) at high speed and high load of the internal combustion engine 1d which is a diesel engine, and even if it is a spark ignition type internal combustion engine, except for FIG. The same is true for the timing chart, although the scavenging timing slightly fluctuates due to the supercharging pressure and the like, but there is no big difference.
図１４は、前記実施例４（図９）の掃気増幅手段を設けた内燃機関の高速回転時の試算によるＰＶ線図である。
図１４のＰＶ線図の縦軸は筒内圧力Ｐで、前記図１３のシリンダ４１ｄの筒内圧力と同じ絶対圧力（ａｂｓ）であり、横軸は前記内燃機関１ｄの前記ＢＤＣとＴＤＣの間のピストン移動による行程容積Ｖｓｔである。
図１４は、ディーゼル機関である前記内燃機関１ｄの試算による複合サイクルのＰＶ線図であるが、火花点火式内燃機関の場合とは、燃焼サイクルが異なるが、排気行程と掃気行程の筒内圧力の値は異なるが挙動は同じである。
図中の太線（点Ｅ４～点Ｅ３）は、前記内燃機関１ｄの高速高負荷運転時のＰＶ線図で、排気行程は、排気弁４７ｄが開弁する点Ｅ３からピストン４２ｄがＢＤＣを点Ｅ４で折り返してピストン４２ｄがＴＤＣ側に移動中に排気弁４７ｄが閉弁する点ＥＳ６までであり、圧縮行程は、前記点ＥＳ６から点Ｓ７を通ってＴＤＣの点Ｃ１までであり、燃焼行程は、前記点Ｃ１でＴＤＣを折り返して点Ｂ２を通って前記点Ｅ３までである。
掃気行程は、排気行程中の点ＥＳ５から前記排気弁４７ｄが閉弁する点ＥＳ６を通って圧縮行程の点Ｓ７までである。
図中の点Ｂ２から点Ｅ３、点Ｓ７から点Ｃ１、および想像線である２点鎖線の点ＥＳ６から点Ｐ１ｄは、理想気体の状態方程式（ＰＶ＝ｎＲＴ）より求めた断熱膨張（Ｂ２～Ｅ３）あるいは断熱圧縮（Ｓ７～Ｃ１）（ＥＳ６～Ｐ１ｄ）の状態変化である。
掃気行程において、排気弁４７ｄの閉弁後に掃気が吸気弁４６ｄからシリンダ４２ｄに流入することにより、点ＥＳ６から点Ｓ７に筒内圧力が上昇し、（Ｎａ）で示す自然吸気時の断熱圧縮（ＥＳ６～Ｐ１ｄ）から、過給時の断熱圧縮（Ｓ７～Ｃ１）の２サイクル内燃機関の燃焼サイクルとなるので、圧縮仕事が増大し内燃機関１ｄの外部に対する仕事は減少するが、燃焼行程で燃料と吸気の燃焼反応によりシリンダ４２ｄ内の吸気の体積が増大するので、過給による前記圧縮仕事の増大により外部に対する仕事の減少の数倍の膨張仕事の増大が発生し、内燃機関１ｄの外部に対する仕事Ｗ（太線内のハッチング部）が過給により増大する。
以上のように、掃気増幅手段５による過給は、内燃機関１ｄの出力が増大するので、内燃機関のダウンサイジングができる。 FIG. 14 is a PV diagram based on a trial calculation at high speed rotation of an internal combustion engine provided with the scavenging air amplification means of the fourth embodiment (FIG. 9).
The vertical axis of the PV diagram of FIG. 14 is the in-cylinder pressure P, which is the same absolute pressure (abs) as the in-cylinder pressure of the cylinder 41 d of FIG. 13. The horizontal axis is between the BDC and TDC of the internal combustion engine 1d. Stroke volume Vst due to piston movement of the
FIG. 14 is a PV diagram of a combined cycle based on a trial calculation of the internal combustion engine 1 d which is a diesel engine, but the combustion cycle is different from that of the spark ignition type internal combustion engine. Although the value of is different, the behavior is the same.
The thick lines (points E4 to E3) in the figure are PV diagrams of the internal combustion engine 1d during high speed and high load operation, and the exhaust stroke is from point E3 at which the exhaust valve 47d opens to piston E 42d at point E4. And the compression stroke is from the point ES6 through the point S7 to the point C1 of the TDC while the piston 42d is moving to the TDC side, and the combustion stroke is At the point C1, the TDC is turned back to pass through the point B2 to the point E3.
The scavenging stroke is from the point ES5 in the exhaust stroke to the point S7 of the compression stroke through the point ES6 at which the exhaust valve 47d closes.
The adiabatic expansion (B2 to E3) obtained from the point B2 to the point E3 in the figure, the point S7 to the point C1, and the point ES6 to the point P1d of the dashed double-dotted line that is an imaginary line Or adiabatic compression (S7 to C1) (ES6 to P1d).
In the scavenging stroke, scavenging flows into the cylinder 42d from the intake valve 46d after the exhaust valve 47d is closed, so that the pressure in the cylinder rises from the point ES6 to the point S7, and adiabatic compression at natural intake shown by (Na) From ES 6 to P 1 d), since the combustion cycle of the two-stroke internal combustion engine with adiabatic compression (S 7 to C 1) at supercharging is achieved, the compression work increases and the work for the outside of the internal combustion engine 1 d decreases, Since the volume of intake air in the cylinder 42d is increased due to the combustion reaction between the engine and the intake air, the increase in the compression work due to supercharging causes an increase in expansion work several times as much as the decrease in the work to the outside, to the outside of the internal combustion engine 1d. Work W (hatched part in thick line) increases due to supercharging.
As described above, since the output of the internal combustion engine 1d is increased by the supercharging by the scavenging air amplification means 5, downsizing of the internal combustion engine can be performed.
図１５は、実施例５（請求項１対応）の、リフト逆止弁と従来技術（特開２０１６－１２５４２１）のトランスベクタを設けた掃気増幅手段と往復圧縮機を設けた内燃機関の断面図である。
図１５は、掃気供給手段として、内燃機関１ｕにて駆動する圧縮機である往復圧縮機２５ｕと、吸気通路である吸気流入通路２２ｕと吸気流出通路２３ｕの間に掃気増幅手段５ｕを設け、前記掃気増幅手段５ｕはリフト逆止弁５５５ｕと前記リフト逆止弁５５５ｕの下流に設けた空気流量増幅器である従来技術（特開２０１６－１２５４２１）のトランスベクタ５３ｕから成り、前記空気流量増幅器であるトランスベクタ５３ｕの駆動流通路５８ｕを前記往復圧縮機２５ｕの吐出口である吐出弁２５７ｕに連通する請求項１に記載の２サイクル内燃機関１ｕである。
前記駆動流通路５８ｕは冷却フィンを設けて断熱圧縮した駆動流を空冷しているが、往復圧縮機２５ｕ等の液冷冷却と共用することもできる。
吸気弁４６ｕは逆止弁で吸気流通路２３ｕと燃焼室の圧力差により、排気弁４７ｕは図示しない容積型油圧供給手段８ｕで発生する油圧を油圧通路８８ｕから弁シリンダ４７１ｕに供給して弁の開閉作動を行う。
往復圧縮機２５ｕの連結棒２５３ｕは位相を進めてストロークを短縮するために出力手段の連結棒４３ｕにピギーバック接続し、往復圧縮機２５ｕのシリンダ２５１ｕは出力手段のシリンダ４１ｕより小さくし、跳ねかけ潤滑を共用できる。
排気弁４７ｕの弁駆動は前記実施例１と同じで、往復圧縮機２５ｕと掃気増幅手段５ｕの作用は前記実施例４と同じで説明を省略し、掃気増幅手段５ｕの構成と作用は、図１６にて説明する。 FIG. 15 is a cross-sectional view of an internal combustion engine provided with scavenging amplification means provided with a lift check valve according to the fifth embodiment (corresponding to claim 1) and a transvector according to the prior art (Japanese Patent Laid-Open No. 2016-125421) and a reciprocating compressor. It is.
FIG. 15 shows, as scavenging air supply means, a reciprocating compressor 25u which is a compressor driven by the internal combustion engine 1u, and a scavenging air amplification means 5u provided between an intake air inflow passage 22u and an intake air outflow passage 23u which are intake air passages. The scavenging air amplification means 5u comprises a lift check valve 555u and a transformer vector 53u of the prior art (Japanese Patent Laid-Open No. 2016-125421) which is an air flow amplifier provided downstream of the lift check valve 555u. The two-stroke internal combustion engine 1u according to claim 1, wherein the drive flow passage 58u of the vector 53u is in communication with a discharge valve 257u which is a discharge port of the reciprocating compressor 25u.
The drive flow passage 58u is provided with cooling fins to air-cool the adiabaticly compressed drive flow, but may be shared with liquid cooling and cooling such as a reciprocating compressor 25u.
The intake valve 46u is a check valve, and the exhaust valve 47u supplies hydraulic pressure generated by a positive displacement hydraulic pressure supply means 8u (not shown) from the hydraulic passage 88u to the valve cylinder 471u by the pressure difference between the intake flow passage 23u and the combustion chamber. Open and close.
The connecting rod 253u of the reciprocating compressor 25u is piggyback connected to the connecting rod 43u of the output means to advance the phase and shorten the stroke, and the cylinder 251u of the reciprocating compressor 25u is smaller than the cylinder 41u of the output means and splashed It can share lubrication.
The valve drive of the exhaust valve 47u is the same as that of the first embodiment, and the operation of the reciprocating compressor 25u and the scavenging amplification means 5u is the same as that of the fourth embodiment and the explanation is omitted. This will be described in 16.
図１６は、前記実施例５（図１５）の従来技術（特開２０１６－１２５４２１）の可変ノズル型のトランスベクタとリフト逆止弁で構成される掃気増幅手段５ｕの断面図である。図１６の前記トランスベクタ５３ｕは、ノズル５３１の上流側のノズル面を設けたピストン５３４を、ノズルが閉じる方向に付勢するスプリング５３５から成るノズル調整機構をハウジング５３３の内周面に設け、前記ノズル調整機構のピストン５３４とフランジ５３６の間にディスク５５７とスプリング５５６からなるリフト逆止弁５５５ｕを設ける。
前記リフト逆止弁５５５ｕは、フランジ５３６に設けたシリンダ部に、ディスク５５７と該ディスク５５７をシリンダ部の座面にスプリング５５６で付勢し、前記ディスク５５７には前記座面に着座時に閉鎖される複数の連通口と、略リング状の外周端部にストロークを規制する当たりと、中央部に吸気流れを円滑にして通路抵抗を小さくするガイド凸部を設けている。 FIG. 16 is a cross-sectional view of the scavenging amplification means 5u configured of a variable nozzle type transvector and a lift check valve according to the prior art (Japanese Patent Laid-Open No. 2016-125421) of the fifth embodiment (FIG. 15). The transformer vector 53u of FIG. 16 is provided with a nozzle adjustment mechanism consisting of a spring 535 for urging the piston 534 provided with the nozzle surface on the upstream side of the nozzle 531 in the nozzle closing direction on the inner peripheral surface of the housing 533 A lift check valve 555 u consisting of a disk 557 and a spring 556 is provided between the piston 534 and the flange 536 of the nozzle adjustment mechanism.
The lift check valve 555 u urges the disc 557 and the disc 557 to a bearing surface of the cylinder portion by a spring 556 against a cylinder portion provided on the flange 536, and the disc 557 is closed when seated on the seat surface. A plurality of communication ports, a contact for restricting the stroke at the substantially ring-shaped outer peripheral end, and a guide convex portion at the central portion for smoothing the intake flow and reducing the passage resistance are provided.
掃気増幅手段５ｕの作用は、リフト逆止弁５５５ｕにて下流の吸気の圧力の上昇により発生する吸気の逆流時に、スプリング５５６の付勢力と逆流吸気によりディスク５５７がフランジ５３６の座面に付勢されて、該トランスベクタ５３ｕによる逆流流量増幅現象を防止し、該リフト逆止弁５５５ｕの下流の圧力が上流より低くなるとリフト逆止弁５５５ｕを開弁して吸気流入通路２２ｕから供給される吸気を空気流出通路２３ｕに流出する。
前記ノズル５３１噴射部の通路径は、前後の吸気流入通路２２ｕと吸気流出通路２３ｕより大きいので、前記ノズル５３１噴射部では吸気は減速し、その減速した吸気を駆動流で流量増幅するので、効率よく流量増幅でき、高速運転にも対応できる。
掃気増幅手段５ｕは、トランスベクタ５３ｕのノズル調整機構により駆動流の圧力と掃気の流量状況に対応した駆動流の流量制御にて掃気の流量増幅を行い、リフト逆止弁５５５ｕにより逆流流量増幅現象を防止して前記実施例４の図１１の（Ｓ３）に示すように駆動流で直接過給を行う。 The action of the scavenging air amplification means 5u is that the disk 557 is urged to the bearing surface of the flange 536 by the urging force of the spring 556 and the backflow air intake at the time of the backflow of the air intake generated by the pressure increase of the intake air downstream by the lift check valve 555u. To prevent the reverse flow rate amplification phenomenon by the transformer vector 53u, and when the pressure downstream of the lift check valve 555u becomes lower than the upstream, the lift check valve 555u is opened and the intake air supplied from the intake inflow passage 22u To the air outflow passage 23 u.
Since the passage diameter of the nozzle 531 injection part is larger than the front and rear intake inflow passages 22 u and the intake outflow passage 23 u, in the nozzle 531 injection part, the intake is decelerated and the decelerated intake is amplified by the drive flow, so the efficiency The flow rate can be well amplified, and it can cope with high speed operation.
The scavenging air amplification means 5u amplifies the scavenging air flow rate by controlling the flow rate of the driving flow according to the driving flow pressure and scavenging flow rate conditions by the nozzle adjustment mechanism of the transformer vector 53u, and the reverse flow rate amplification phenomenon by the lift check valve 555u. As shown in (S3) of FIG. 11 of the fourth embodiment, direct supercharging is performed by the drive flow.
図１７は、実施例６（請求項１対応）の掃気増幅手段と往復圧縮機を設けた２気筒内燃機関１ｋの構成説明図の平面図（１）、Ｋ－Ｋ断面図（２）、Ｌ－Ｌ断面図（３）である。図１７は、上図（１）の２気筒内燃機関１ｋの平面図に示すように、燃焼室が略球面状の各気筒に往復圧縮機２５（ｋ、Ｌ）を設け、掃気増幅手段５ｋに前記往復圧縮機２５（ｋ、Ｌ）からの駆動流を駆動流通路５８ｋから供給し、掃気増幅手段５ｋから各気筒のタンゼンシャルポートに掃気を流量増幅して供給して内燃機関１ｋの掃気と過給を行う。
中図（２）のＫ－Ｋ断面と、下図（３）のＬ－Ｌ断面に示すように、各気筒のクランク角は１８０°異なるので、図１８に示すように掃気の作動タイミングは干渉しないので、各往復圧縮機２５（ｋ、Ｌ）と掃気増幅手段５ｋの作用は、前記実施例４と同じである。
掃気増幅手段５ｋを共用できるので、簡素な構成で安価に製作でき、燃焼効率が高く高出力の内燃機関１ｋである。 FIG. 17 is a plan view (1), a sectional view taken along the line KK (2), L of a structural explanatory view of a two-cylinder internal combustion engine 1k provided with the scavenging air amplification means of the sixth embodiment (corresponding to claim 1) and a reciprocating compressor. FIG. 6 is a cross-sectional view (3). As shown in the plan view of the two-cylinder internal combustion engine 1k in the upper diagram (1) in FIG. 17, a reciprocating compressor 25 (k, L) is provided in each cylinder whose combustion chamber is substantially spherical, The drive flow from the reciprocating compressor 25 (k, L) is supplied from the drive flow passage 58k, and the scavenging air is amplified from the scavenging amplification means 5k to the tangential port of each cylinder and supplied to the scavenging of the internal combustion engine 1k. And supercharge.
As shown in the middle section (2) K-K section and the lower section (3) L-L section, the crank angle of each cylinder differs by 180 °, so the scavenging operation timing does not interfere as shown in FIG. Therefore, the operation of each reciprocating compressor 25 (k, L) and the scavenging amplification means 5 k is the same as that of the fourth embodiment.
Since the scavenging air amplification means 5k can be shared, the internal combustion engine 1k can be manufactured inexpensively with a simple configuration, has a high combustion efficiency, and has a high output.
図１８は、前記実施例６（図１７）の内燃機関の各気筒のタイミングチャートと試算による筒内圧力である。
図１８は、各項目には前記２気筒（Ｋ－Ｋ）（Ｌ－Ｌ）のデータを記載しているが、図表の作成方法は前記実施例４（図１３）の「内燃機関１ｄのタイミングチャートと試算による筒内圧力」と同じであるので、図表の作成方法の説明は省略する。
図中の極太実線は第１気筒（Ｋ－Ｋ）、太破線（Ｌ－Ｌ）は第２気筒で、吸気流出通路２３ｋが各気筒の燃焼室と吸気弁（ｋ、Ｌ）の開弁により繋がるのは各気筒の掃気行程であるので、図の作動行程で分かるように各気筒の掃気行程はクランク角が１８０°位相か異なるので干渉せず、同様に各往復圧縮機のピストン２５２（ｋ、Ｌ）も吐出タイミングも干渉しない。 FIG. 18 is a timing chart of each cylinder of the internal combustion engine of the sixth embodiment (FIG. 17) and in-cylinder pressure based on trial calculation.
FIG. 18 describes the data of the two-cylinder (K-K) (L-L) in each item, but the chart creation method is the “timing of the internal combustion engine 1 d of the fourth embodiment (FIG. 13) Since it is the same as the “in-cylinder pressure based on the chart and trial calculation”, the description of the method of creating the chart is omitted.
The thick solid line in the figure is the first cylinder (K-K), the thick broken line (L-L) is the second cylinder, and the intake and outflow passages 23k are open by the combustion chambers of the respective cylinders and the intake valves (k, L). Since it is the scavenging stroke of each cylinder that is linked, the scavenging stroke of each cylinder does not interfere because the crank angle is 180 ° out of phase or not, as can be seen in the operation strokes of the figure. , L) and ejection timing do not interfere.
図１９は、実施例７（請求項１対応）の掃気増幅手段と往復圧縮機を設けた３気筒内燃機関の構成説明図の平面図（１）と前記平面図（１）のＭ－Ｍ断面図（２）である。
内燃機関１ｍは、各気筒の位相差が１２０°の直列３気筒の出力手段の各々の気筒に各々の気筒より位相が進んだ往復圧縮機２５（ｍ１～ｍ３）を並行に設け、出力手段のクランク軸４４ｍと往復圧縮機２５のクランク軸２５４ｍ歯車で連動し、前記全ての往復圧縮機の吐出口を吸気流入通路２２ｍと吸気流出通路２３ｍの間に設けた掃気増幅手段５ｍの空気流量増幅器５０ｍの駆動流通路５８ｍに連通し、前記吸気流出通路２３ｍを前記直列３気筒の全ての吸気弁（タンゼンシャルポート）に連通する。
前記実施例６と同様に、各掃気行程は干渉しないが、往復圧縮機２５の吐出が部分的に緩衝するが大勢には影響がなく、内燃機関１ｍのシリンダを密集できるので、小型で出力の大きな内燃機関にでき、燃焼性等のその他の利点は同じであるので説明を省略する。 FIG. 19 is a plan view (1) of a configuration explanatory view of a three-cylinder internal combustion engine provided with the scavenging air amplification means of the seventh embodiment (corresponding to claim 1) and a reciprocating compressor, and an MM cross section of the plan view (1). It is a figure (2).
The internal combustion engine 1m is provided with reciprocating compressors 25 (m1 to m3) parallel to each cylinder of the in-line three-cylinder output means whose phase difference between the cylinders is 120.degree. An air flow amplifier 50m of scavenging air amplification means 5m provided with a crankshaft 44m and a crankshaft 254m gear of the reciprocating compressor 25 and the discharge ports of all the reciprocating compressors provided between the intake inflow passage 22m and the intake outflow passage 23m The intake flow passage 23m is in communication with all the intake valves (tangential ports) of the in-line three cylinders.
As in the sixth embodiment, although the scavenging strokes do not interfere with each other, the discharge of the reciprocating compressor 25 partially buffers but there is no effect on many, and the cylinders of the internal combustion engine 1m can be compacted. A large internal combustion engine can be used, and the other advantages such as the flammability are the same, so the description will be omitted.
図２０は、実施例８（請求項１対応）の往復圧縮機２５ｓと、掃気増幅手段５ｓに制御弁５６を設けた２サイクル内燃機関１ｓの構成概念の説明図である。
図２０は、掃気供給手段として、前記内燃機関１ｓにて駆動する往復圧縮機２５ｓと、吸気通路５８ｓに掃気増幅手段５ｓとを設け、前記掃気増幅手段５ｓは逆止弁５５ｓと前記逆止弁５５ｓの下流に設けた空気流量増幅器５０ｓから成り、前記空気流量増幅器５０ｓの駆動流通路５８ｓを前記往復圧縮機２５ｓの吐出口に連通する請求項１に記載の２サイクル内燃機関１ｓである。
前記掃気増幅手段５ｓにおいて、前記空気流量増幅器５０ｓの上流に制御弁５６を設け、前記内燃機関１ｓの運転状況に応じて前記制御弁５６を制御し、掃気の圧力および流量を調整する。
図２０は、前記実施例４（図１）の内燃機関１ｄの燃焼室中央に点火プラグ１１ｓを設け、燃焼室に主燃料を供給するインジェクション１２ｓ、駆動流通路５８ｓに冷却器５８１と水素等の空気より密度が小さい気体燃料を供給するインジェクション１２ｓ２を設けた火花点火式内燃機関１ｓである。
排気の性状が良好となる本願発明の内燃機関では、排気通路３１ｓと駆動流通路５８ｓに連通する排気還流通路３６を設け、前記排気還流通路３６に制御弁３８と排気通路３１ｓへの逆流を防止する逆止弁３７を設けて排気圧力を駆動流圧力に利用できるＥＧＲが可能である。 FIG. 20 is an explanatory view of a configuration concept of a two-stroke internal combustion engine 1s in which a control valve 56 is provided in the reciprocating compressor 25s of the eighth embodiment (corresponding to claim 1) and the scavenging amplification means 5s.
In FIG. 20, as the scavenging air supply means, a reciprocating compressor 25s driven by the internal combustion engine 1s and a scavenging air amplification means 5s are provided in the intake passage 58s, and the scavenging air amplification means 5s includes the check valve 55s and the check valve The two-stroke internal combustion engine 1s according to claim 1, comprising an air flow amplifier 50s provided downstream of 55s, wherein the drive flow passage 58s of the air flow amplifier 50s is in communication with the discharge port of the reciprocating compressor 25s.
In the scavenging amplification means 5s, a control valve 56 is provided upstream of the air flow rate amplifier 50s, and the control valve 56 is controlled according to the operating condition of the internal combustion engine 1s to adjust the scavenging pressure and flow rate.
In FIG. 20, an ignition plug 11s is provided at the center of the combustion chamber of the internal combustion engine 1d of the fourth embodiment (FIG. 1) and the main fuel is supplied to the combustion chamber 12s. This is a spark ignition internal combustion engine 1 s provided with an injection 12 s 2 for supplying a gaseous fuel whose density is smaller than that of air.
In the internal combustion engine according to the present invention, the exhaust gas recirculation passage 36 communicating with the exhaust gas passage 31s and the drive flow passage 58s is provided, and backflow to the control valve 38 and the exhaust gas passage 31s is prevented in the exhaust gas recirculation passage 36. It is possible to provide the check valve 37 to enable the EGR that can use the exhaust pressure as the drive flow pressure.
図２０の内燃機関１ｓの作用は、往復圧縮機２５ｓから供給される駆動流で掃気増幅手段５ｓの空気流量増幅器５０ｓにて掃気を流量増幅し、吹き抜けを補填して余りある掃気を供給することにより、完全なガス交換と駆動流による掃気の加圧により過給を行う。
内燃機関１ｓの燃焼室の混合気が過給により過熱するとノッキング等の異常発火するのを防止する効果がある。
制御弁５６の調整により燃焼室のＴＤＣでの吸気温度の制御等によりＨＣＣＩエンジン（予混合圧縮自動着火）またはＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）とすることにより燃焼性の改善と出力の向上の効果がある。
制御弁５６を全開すると掃気増幅手段５ｓは通常の流量増幅による大量の掃気による完全な掃気と過給を行い図２１に示すＰＶ線図の二点鎖線のような燃焼サイクルで外部に対する仕事を行い、制御弁５６の開度を調整することにより流量増幅を抑制して掃気の流量を減少することにより掃気行程での圧力上昇による過給効果を抑制し、前記ＰＶ線図の太線に示すように外部に対する仕事Ｗｂを行う自然吸気運転あるいは吸気を更に削減した運転ができるので内燃機関１ｓの燃焼室の過熱を防止し、燃焼室の吸気の温度を制御できるので、制御弁５６の開度の制御によりＴＤＣでの予混合気の温度をＨＣＣＩ運転ができる範囲に制御でき、制御方法は実施例１３にて説明する。
制御弁５６は、電動式の構造が簡単なバタフライバルブでも、直線的な開度制御ができるポペット式等でもよい。 In the operation of the internal combustion engine 1s of FIG. 20, scavenging air is amplified at the air flow rate amplifier 50s of the scavenging air amplification means 5s by a drive flow supplied from the reciprocating compressor 25s, and scavenging is supplied by compensating for blowby. Thus, supercharging is performed by complete gas exchange and pressurization of scavenging air by drive flow.
When the air-fuel mixture in the combustion chamber of the internal combustion engine 1s is superheated due to supercharging, there is an effect of preventing abnormal ignition such as knocking.
By adjusting the control valve 56 to control the intake air temperature at the TDC of the combustion chamber etc. to improve the combustibility and output by making the HCCI engine (premixed compression auto ignition) or the SPCCI engine (compression ignition combustion by spark control) There is an effect of
When the control valve 56 is fully opened, the scavenging amplification means 5s performs complete scavenging and supercharging with a large amount of scavenging by normal flow rate amplification, and performs work for the outside in a combustion cycle like a two-dot chain line in the PV diagram shown in FIG. By suppressing the flow amplification by adjusting the opening degree of the control valve 56 and reducing the flow rate of scavenging, the supercharging effect due to the pressure increase in the scavenging stroke is suppressed, as shown by the thick line of the PV diagram. Since natural intake operation to perform work Wb to the outside or operation with reduced intake can be further prevented, overheating of the combustion chamber of the internal combustion engine 1s can be prevented, and temperature of intake of the combustion chamber can be controlled. Thus, the temperature of the premixed gas at TDC can be controlled to a range where HCCI operation can be performed, and a control method will be described in Example 13.
The control valve 56 may be a butterfly valve having a simple motorized structure, or may be a poppet type capable of linear opening control.
図２１は、前記実施例８（図２０）の掃気増幅手段に設けた制御弁５６による掃気の流量増幅抑制時の試算によるＰＶ線図である。
図２１の作図方法は、前記実施例４（図１４）と同じであるので説明を省略する。
排気増幅手段５ｓの制御弁５６により掃気と過給の調整ができ、図２１の太線（点Ｅ４ｂ～点Ｅ３ｂ）は、前記内燃機関１ｓの低負荷運転時に前記制御弁５６の制御により過給作用を抑制した自然吸気運転時のＰＶ線図で外部に対する仕事Ｗｂ（太線内のハッチング部）を燃焼サイクル毎に行う。
図２１の二点鎖線は、前記制御弁５６を全開した過給運転時のＰＶ線図で、排気行程と掃気行程の掃気増幅手段の作用は、前記実施例４（図１４）の太線（点Ｅ４～点Ｅ３）の燃焼サイクルと同じであるので、説明を省略する。
掃気増幅手段５による過給は、内燃機関１ｓの出力が増大するので、内燃機関のダウンサイジングができ、更に、掃気増幅手段５に設けた制御弁５６により、内燃機関１ｓの運転状況に対応した掃気と過給によりＴＤＣでの予混合気の温度を調整できるので、ＨＣＣＩエンジンまたはＳＰＣＣＩエンジンへの運転切換ができる。
前記実施例４（図１４）の内燃機関１ｄはディーゼル機関であり、図２１に示す実施例８の内燃機関１ｓは火花点火式内燃機関であり、理論サイクル（サバティサイクルとオットーサイクル）が異なるが、両方の実施例は共に内燃機関の種類を限定するものではなく、ディーゼル機関でも火花点火式内燃機関でもよい。 FIG. 21 is a PV diagram based on a trial calculation at the time of suppression of the flow rate amplification of scavenging by the control valve 56 provided in the scavenging amplification means of the eighth embodiment (FIG. 20).
The drawing method of FIG. 21 is the same as that of the fourth embodiment (FIG. 14), so the description will be omitted.
Scavenging and supercharging can be adjusted by the control valve 56 of the exhaust amplification means 5s, and the thick lines (point E4b to point E3b) in FIG. 21 indicate supercharging action by control of the control valve 56 during low load operation of the internal combustion engine 1s. Work Wb (hatched part in thick line) to the outside is performed for each combustion cycle in the PV diagram during natural intake operation with suppressed.
The two-dot chain line in FIG. 21 is a PV diagram at the time of supercharging operation with the control valve 56 fully opened, and the function of the scavenging air amplification means of the exhaust stroke and scavenging stroke is the thick line (dots) of the fourth embodiment (FIG. Since this is the same as the combustion cycle at E4 to E3), the description will be omitted.
The supercharging by the scavenging air amplification means 5 increases the output of the internal combustion engine 1s, so downsizing of the internal combustion engine can be performed, and the control valve 56 provided in the scavenging air amplification means 5 corresponds to the operating condition of the internal combustion engine 1s. Since the temperature of the premixed gas at TDC can be adjusted by scavenging and supercharging, the operation can be switched to an HCCI engine or an SPCCI engine.
The internal combustion engine 1d according to the fourth embodiment (FIG. 14) is a diesel engine, and the internal combustion engine 1s according to the eighth embodiment shown in FIG. 21 is a spark ignition internal combustion engine. The theoretical cycles (Sabati cycle and Otto cycle) are different. However, both embodiments do not limit the type of internal combustion engine, and may be a diesel engine or a spark ignition internal combustion engine.
図２２は、実施例９（請求項２対応）の燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、排気弁を２本の各カム軸に設けたカムにより開閉し、吸気弁を逆止弁とする掃気増幅手段を設けた２気筒２サイクル内燃機関の構成概念の説明図である。
図２２は、燃焼室４１０ｎに放射状に複数の吸気弁４６ｎと排気弁４７ｎを交互に配置し、クランク軸４４ｎの回転数と同じ回転数で連動する平行な２本のカム軸４０７（－１、－２）を設け、前記排気弁４７ｎを前記２本の各カム軸４０７（－１、－２）に設けたカム４０８（－１、－２）により開閉し、前記吸気弁４６ｎを逆止弁であるリフト逆止弁とする請求項１に記載の２サイクル内燃機関１ｎである。
吸気弁４６ｎは、リフト逆止弁で、掃気と燃焼室の圧力差により開弁する。
内燃機関１ｎは、クランク軸４４ｎに設けた駆動車４０１で伝動媒体４０３ｎを介してカム軸４０７-１に設けた駆動車４０１と同じ有効径（φＤｎ）の従動車４０２を駆動し、カム軸４０７－１をクランク軸４４ｎに同期回転する。
前記カム軸４０７－１に設けた駆動歯車４０５に噛合うカム軸４０７－２に設けた前記駆動歯車４０５とピッチ円直径が同じ従動歯車４０６により、カム４０８（－１、－３）を設けたカム軸４０７－１とカム４０８（－２、－４）を設けたカム軸４０７－２は同一回転数で逆方向に回転する。
カム４０８―１とカム４０８－２は、シリンダ軸に対し左右対称のカム形状とし、それぞれのカムで作動する各排気弁４７ｎはクランク軸４４ｎに同期して開弁する。
燃焼室４１０ｎに放射状に複数の吸気弁４６ｎと排気弁４７ｎを交互に配置し、更に吸気弁４６ｎ同士と排気弁４７ｎ同士は同一線上に配置し、下図に示すように吸気弁４６ｎと排気弁４７ｎの狭角をθｎ（θｎ>９０°）とすることにより、多気筒の弁の配置干渉を軽減し、気筒間の距離を短縮することができ、小型軽量で、剛性の大きいシリンダブロックにできる。
内燃機関１ｎのタンゼンシャルポート２３０ｎとインジェクタ１２ｎから供給される燃料の作用は実施例１と、往復圧縮機２５ｎと掃気増幅手段５ｎの作用は実施例８と重複するので説明を省略する。 In FIG. 22, a plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber of the ninth embodiment (corresponding to claim 2), and exhaust valves are opened and closed by cams provided on each of two camshafts. It is explanatory drawing of the construction concept of the 2 cylinder 2 cycle internal combustion engine which provided the scavenging air amplification means which makes a valve a non-return valve.
In FIG. 22, a plurality of intake valves 46n and exhaust valves 47n are alternately arranged radially in the combustion chamber 410n, and two parallel camshafts 407 (-1, -1, interlocked at the same rotational speed as the rotational speed of the crankshaft 44n). 2), and the exhaust valve 47n is opened and closed by a cam 408 (-1, 2) provided on each of the two camshafts 407 (-1, 2), and the intake valve 46n is A two-stroke internal combustion engine 1n according to claim 1, wherein the lift check valve is
The intake valve 46 n is a lift check valve, and is opened by the pressure difference between the scavenging air and the combustion chamber.
The internal combustion engine 1n drives a driven vehicle 402 provided on the crankshaft 44n via a transmission medium 403n to drive a driven vehicle 402 having the same effective diameter (φ Dn) as the drive vehicle 401 provided on the camshaft 407-1. Synchronize the rotation of -1 to the crankshaft 44n.
A cam 408 (-1, -3) is provided by a driven gear 406 having the same pitch circle diameter as the drive gear 405 provided on the cam shaft 407-2 meshing with the drive gear 405 provided on the cam shaft 407-1. The cam shaft 407-1 provided with the cam shaft 407-1 and the cam 408 (-2, 4) rotates in the opposite direction at the same rotation speed.
The cams 408-1 and 408-2 have cam shapes that are symmetrical with respect to the cylinder axis, and the exhaust valves 47n operated by the respective cams open in synchronization with the crankshaft 44n.
A plurality of intake valves 46n and exhaust valves 47n are alternately arranged radially in the combustion chamber 410n, and the intake valves 46n and the exhaust valves 47n are disposed on the same line, and the intake valves 46n and the exhaust valves 47n as shown in the figure below. By setting the narrow angle of θ to θn (θn> 90 °), the arrangement interference of multi-cylinder valves can be reduced, the distance between cylinders can be shortened, and a compact, lightweight, rigid cylinder block can be obtained.
The operation of the fuel supplied from the tangential port 230n and the injector 12n of the internal combustion engine 1n is the same as that of the first embodiment, and the operation of the reciprocating compressor 25n and the scavenging amplification means 5n is the same as that of the eighth embodiment.
図２３は、実施例１０（請求項２対応）の２本のカム軸により、排気弁がカム駆動で吸気弁が油圧手段を介して油圧駆動する２気筒２サイクル内燃機関の平面図と周辺回路図である。
図２３は、燃焼室に放射状に複数の吸気弁４６ｐと排気弁４７ｐを交互に配置し、クランク軸４４ｐに駆動車４０１ｐ、カム軸４０７ｐに前記駆動車４０１ｐと同じ有効径の従動車４０２ｐと駆動歯車４０５ｐを設け、カム軸４０７ｐ２に前記駆動歯車４０５ｐに噛合う同じピッチ円直径の従動歯車４０６ｐを設け、前記クランク軸４４ｐの回転数と同じ回転数で連動する平行な２本の前記カム軸（４０７ｐ、４０７ｐ２）を設け、排気弁４７ｐを前記２本の各カム軸（４０７ｐ、４０７ｐ２）に設けたカム４０８ｐにより開閉し、前記吸気弁４６ｐを前記カムとは別のカムに連動する油圧手段である弁駆動ユニット８０（ｐ１、ｐ２、ｐ３）により開閉する請求項１に記載の２サイクル内燃機関１ｐである。
前記駆動車４０１ｐは伝動媒体４０３ｐを介して従動車４０２ｐを駆動し、前記弁駆動ユニット８０（ｐ１、ｐ２、ｐ３）はカムに連動するプランジャ８４（ｐ１、ｐ２、ｐ３）で発生する油圧にて吸気弁４６ｐを開弁する。
内燃機関１ｐのタンゼンシャルポート２３０ｎとインジェクタ１ｐから供給される燃料の作用は実施例１と、往復圧縮機２５ｐと掃気増幅手段５ｐの作用は実施例８と重複するので説明を省略する。 FIG. 23 is a plan view and peripheral circuits of a two-cylinder two-stroke internal combustion engine in which the exhaust valve is cam driven and the intake valve is hydraulically driven via hydraulic means by two camshafts of the tenth embodiment (corresponding to claim 2) FIG.
In FIG. 23, a plurality of intake valves 46p and exhaust valves 47p are alternately arranged radially in the combustion chamber, the drive wheel 401p is for the crankshaft 44p, and the driven wheel 402p with the same effective diameter as the drive wheel 401p is for the camshaft 407p. A gear 405p is provided, and the cam shaft 407p2 is provided with a driven gear 406p having the same pitch circle diameter that meshes with the drive gear 405p, and two parallel cam shafts interlocked at the same rotation speed as the rotation speed of the crankshaft 44p ( 407p, 407p2), and the exhaust valve 47p is opened and closed by a cam 408p provided on each of the two cam shafts (407p, 407p2), and the intake valve 46p is operated by hydraulic means interlocking with a cam other than the cam The two-stroke internal combustion engine 1p according to claim 1, which is opened and closed by a valve drive unit 80 (p1, p2, p3).
The driving wheel 401p drives the driven wheel 402p via the transmission medium 403p, and the valve driving unit 80 (p1, p2, p3) is a hydraulic pressure generated by the plunger 84 (p1, p2, p3) interlocking with the cam. Open the intake valve 46p.
The actions of the fuel supplied from the tangential port 230n of the internal combustion engine 1p and the injector 1p are the same as those of the first embodiment, and the actions of the reciprocating compressor 25p and the scavenging amplifier 5p are the same as those of the eighth embodiment.
図２４は、前記実施例１０（図２３）の、ＴＤＣの各燃料濃度層の分布状況と水素可燃層の点火時のＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）の説明図である。
図２４は、前記実施１０（図２３）のピストン頂面の中央に半径ＳＲｐ２の球面状キャビティ４２０ｐを設けたピストン４７ｇがＴＤＣにて点火プラグ１１ｐにて点火した状態の説明図で、吸気弁４６ｐは弁駆動ユニット８０ｐ２から供給される油圧により開弁し、排気弁４７ｐはカム軸４０７ｐに設けたカム４０８ｐにより開弁する。
ＴＤＣでは全ての弁が閉弁し、断熱圧縮により予混合吸気はガソリンの発火点（３００℃）以下の温度まで昇温し、エジェクタ１２ｐ１から供給された水素は前記実施例２で説明したスワールの遠心分離作用により高濃度層Ｆ１ｐから順次水素の濃度が低下する同心円状の分布層（Ｆ２ｐ～Ｆ４ｐ）を形成し、インジェクタ１２ｐ２から供給された主燃料であるガソリンは、密度が大きいので低濃度層Ｆ３ｐ付近に拡散する。
点火プラグ１１ｐにて点火した水素は、高濃度層Ｆ１ｐから高速火炎伝播して燃焼し、高濃度層側Ｆ１ｐに拡散した一部のガソリンと一緒に高速燃焼するので、燃焼室の圧力上昇により温度は急激に上昇してガソリンの発火点を超えるので燃焼室内で一斉にガソリンが燃焼し、燃焼が進むにつれ更に水素の発火点（５８５℃）も超えるが、その時点では水素の高速火炎伝播により既に水素は殆んど燃焼している。
このように、掃気増幅手段５ｐの制御弁５６ｐの制御により点火プラグ１１ｐの火花点火により高速燃焼が可能なＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）とし、燃焼性を改善し、内燃機関１ｐの熱効率の向上と出力の増大ができる。
前記ＴＤＣでの予混合気の温度を制御弁５６ｐの制御によりガソリンの発火点（３００℃）以上に制御することによりＨＣＣＩエンジン（予混合圧縮自動着火）とすることもできる。
図２４に示すように、上記作用によりピストン頂面の球面状キャビティ４２０ｐには燃料濃度の低い複数の環状層が主に接触するので、ピストン４２ｐにより高濃度可燃層（Ｆ１ｐ）等の燃焼が妨げられないので燃焼性が向上し、ピストン４２ｐの頂面の大半の面積が低い濃度層（Ｆ３ｐ、Ｆ４ｐ）に接するので熱損失が抑制されて出力が向上し、ピストン４２ｐの過熱を抑制する効果がある。
スワールによる遠心分離作用等の作用は前記実施例２と重複するので説明を省略する。 FIG. 24 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the tenth embodiment (FIG. 23).
FIG. 24 is an explanatory view of a state where a piston 47g provided with a spherical cavity 420p of radius SRp2 at the center of the piston top surface in the above-mentioned Embodiment 10 (FIG. 23) is ignited by the spark plug 11p at TDC. The valve is opened by the hydraulic pressure supplied from the valve drive unit 80p2, and the exhaust valve 47p is opened by a cam 408p provided on the cam shaft 407p.
In TDC, all the valves are closed, the adiabatic compression raises the temperature of the premixed intake air to the temperature below the ignition point (300 ° C.) of gasoline, and the hydrogen supplied from the ejector 12p1 is the swirl described in the second embodiment. Concentrated distribution layers (F2p to F4p) in which the concentration of hydrogen decreases sequentially from the high concentration layer F1p by the centrifugal separation action are formed, and gasoline, which is the main fuel supplied from the injector 12p2, has a large density, so the low concentration layer Diffuse around F3p.
The hydrogen ignited by the spark plug 11p propagates at a high speed from the high concentration layer F1p and burns, and burns together with a part of the gasoline diffused to the high concentration layer side F1p. The temperature rises rapidly and exceeds the ignition point of gasoline, so gasoline is burned all at once in the combustion chamber, and the ignition point (585 ° C) of hydrogen is further exceeded as the combustion proceeds, but at that point the high speed flame propagation of hydrogen already Hydrogen is almost burning.
Thus, by controlling the control valve 56p of the scavenging air amplification means 5p, the SPCCI engine (compressive ignition combustion by spark control) is made possible by high-speed combustion by spark ignition of the spark plug 11p, and the combustibility is improved, and the thermal efficiency of the internal combustion engine 1p Improve output and increase output.
The HCCI engine (premixed compression auto-ignition) can also be obtained by controlling the temperature of the premixed gas at TDC above the ignition point (300 ° C.) of gasoline by controlling the control valve 56p.
As shown in FIG. 24, the annular thin layer with low fuel concentration is mainly in contact with the spherical cavity 420p on the top face of the piston by the above action, and the combustion of the high concentration combustible layer (F1p) is hindered by the piston 42p. Combustion is improved and most of the area on the top surface of the piston 42p is in contact with the low concentration layer (F3p, F4p), so the heat loss is suppressed and the output is improved. is there.
The action of the centrifugal separation action and the like by the swirl is the same as that of the second embodiment, so the description will be omitted.
図２５は、実施例１１（請求項３対応）の掃気増幅手段と酸水素発生装置を備えた水素を燃料とする内燃機関の構成概念の説明図である。
図２５は、内燃機関１ｈに電気的手段により運転する酸水素発生装置９を設け、前記酸水素発生装置９で発生する酸水素を、前記空気より密度が小さい燃料として供給する請求項２に記載の２サイクル内燃機関１ｈである。
流体供給手段７ｈの燃料タンク７５ｈに加圧貯蔵された水素を流体制御手段６ｈの減圧弁６４ｈで減圧し、酸水素発生手段９の酸水素発生装置９０を二次電池９６で作動して電解液タンク９４の電解液を電気分解して発生する酸水素を流体制御手段６ｈの制御弁６３ｈで供給量を調整し、前記水素と酸水素を高圧燃料ポンプ１３ｈに供給して加圧し、インジェクタ１２ｈから適時燃焼室に燃料噴射する。水素と酸水素の混合比は、燃料センサ６２ｈで調整し、燃料噴射量を決定する。
二次電池９６の替わりに、内燃機関１ｈにより駆動される発電機により発生する電力により前記酸水素発生装置９０を運転することもできる。
電気分解で酸水素を発生できるので、水素の自給または燃料タンク７５ｈの水素の補充量を減少でき、酸水素発生装置９０の二次電池９６を利用して実施例１３に示すようなハイブリッド車両のエネルギ回生ができる。
排気弁４７ｈの弁駆動機構、往復圧縮機２５ｈ等の構成、作用は実施例９と同じであるので説明を省略する。 FIG. 25 is an explanatory view of a configuration concept of an internal combustion engine using hydrogen as a fuel and provided with the scavenging air amplification means and the oxyhydrogen generator according to Embodiment 11 (corresponding to claim 3).
FIG. 25 shows the internal combustion engine 1h provided with an oxyhydrogen generator 9 operated by electrical means, and supplying the oxyhydrogen generated by the oxyhydrogen generator 9 as a fuel having a density smaller than that of the air. 2 cycle internal combustion engine 1h.
The hydrogen stored under pressure in the fuel tank 75h of the fluid supply means 7h is depressurized by the pressure reducing valve 64h of the fluid control means 6h, and the oxyhydrogen generation device 90 of the oxyhydrogen generation means 9 is operated with the secondary battery 96 to make the electrolyte The amount of oxygen and hydrogen generated by electrolyzing the electrolytic solution in the tank 94 is adjusted by the control valve 63h of the fluid control means 6h, and the hydrogen and the acid hydrogen are supplied to the high pressure fuel pump 13h to pressurize them. Fuel injection to the combustion chamber at appropriate times. The mixing ratio of hydrogen and oxyhydrogen is adjusted by the fuel sensor 62h to determine the fuel injection amount.
Instead of the secondary battery 96, the oxyhydrogen generator 90 can also be operated by the electric power generated by the generator driven by the internal combustion engine 1h.
Since oxyhydrogen can be generated by electrolysis, it is possible to reduce the self-supply of hydrogen or the replenishment amount of hydrogen in the fuel tank 75h, and the hybrid vehicle as shown in Example 13 using the secondary battery 96 of the oxyhydrogen generator 90. Energy regeneration is possible.
The structure of the valve drive mechanism of the exhaust valve 47h, the configuration of the reciprocating compressor 25h, and the like, and the operation are the same as those of the ninth embodiment, and thus the description thereof will be omitted.
図２６は、実施例１２（請求項３対応）の、従来技術（特許文献９）の電解液に超音波を付加する酸水素発生装置の構成概念の説明図で、前記実施例１１（図２５）の前記酸水素発生装置の一例である。
図２６は、電解槽９１４内に層状配置する陰極９１１と陽極９１２と、前記陰極９１１と前記陽極９１２との間に直流電圧を印加する直流電源９１３と、電解液９１５の供給を制御する電解液制御手段であるポンプ９１７と、電気分解にて発生する気体を捕集する気体捕集手段である制御弁９１８と、で構成する電気分解手段９１と、超音波発振子９２１と、前記超音波発振子９２１を電気的手段により超音波振動させる高周波発生器９２２と、で構成する超音波発振手段９２と、を備えた酸水素発生装置９０ｒにおいて、前記超音波振動子と前記電解液を伝搬する超音波振動の反射面との距離を超音波波長の４分の１の整数倍とし、前記電気分解手段９１は、前記陰極９１１と前記陽極９１２を、前記超音波発振子９２１から発振する超音波振動の伝搬方向に対し垂直な平面上に配置し、超音波振動が前記陰極９１１と前記陽極９１２を通過して伝搬できるように前記陰極９１１と前記陽極９１２をスラット状またはグリッド状とし、更に、前記超音波発振子９２１の振動面から、超音波波長（λ）の4分の１の奇数倍の距離に前記陰極９１１を、超音波波長（λ）の４分の１の偶数倍の距離に前記陽極９１２を、配置する酸水素発生装置６ｒである。
上記酸水素発生装置９０ｒは、電極の配置密度を大きくできるので酸水素発生装置を小型化できる。
更に、電解効率を低下させる陰極に発生する析出物を超音波の剥離除去作用により除去し、陽極に発生する気泡状のガスを超音波の電解液往復運動による気泡脱離作用により陽極から離脱するので、電解効率の低下を防止できる効果があり、移動手段への組み込みが容易となる。 FIG. 26 is an explanatory view of a configuration concept of an oxyhydrogen generator for adding an ultrasonic wave to an electrolytic solution according to the prior art (Patent Document 9) of Example 12 (corresponding to claim 3). And the like.
FIG. 26 shows a cathode 911 and an anode 912 arranged in layers in the electrolytic cell 914, a DC power supply 913 for applying a DC voltage between the cathode 911 and the anode 912, and an electrolyte for controlling the supply of the electrolyte 915. An electrolysis means 91 comprising a pump 917 as control means and a control valve 918 as gas collection means for collecting gas generated by electrolysis, an ultrasonic oscillator 921 and the ultrasonic oscillation In the oxyhydrogen generator 90r provided with an ultrasonic oscillation means 92 comprising: a high frequency generator 922 for ultrasonically vibrating the element 921 by electrical means; and an ultrasonic wave propagating through the ultrasonic vibrator and the electrolytic solution The distance between the acoustic wave and the reflection surface is an integral multiple of a quarter of the ultrasonic wavelength, and the electrolysis means 91 is an ultrasonic wave that oscillates the cathode 911 and the anode 912 from the ultrasonic wave generator 921 The cathode 911 and the anode 912 are slat-like or grid-like so that ultrasonic oscillation can be propagated through the cathode 911 and the anode 912 in a plane perpendicular to the propagation direction of the oscillation. A distance from the vibration plane of the ultrasonic wave generator 921 to an odd multiple of a quarter of the ultrasonic wavelength (λ), a distance from an even multiple of a quarter of the ultrasonic wavelength (λ) This is an oxyhydrogen generator 6r in which the anode 912 is disposed.
Since the above-mentioned oxyhydrogen generator 90r can increase the arrangement density of the electrodes, the oxyhydrogen generator can be miniaturized.
Furthermore, the deposit generated on the cathode which lowers the electrolytic efficiency is removed by the peeling and removing action of the ultrasonic wave, and the bubble-like gas generated on the anode is separated from the anode by the bubble desorbing action due to the electrolyte reciprocation of the ultrasonic wave. Therefore, there is an effect that the reduction of the electrolysis efficiency can be prevented, and the incorporation into the moving means becomes easy.
図２７は、実施例１３（請求項３対応）の掃気増幅手段、酸水素発生装置、と回生手段を設けたハイブリッド車両の火花点火式内燃機関の構成概念の説明図である。
図２７の内燃機関１ｔは、前記実施例８（図２０）の内燃機関１ｓの吸気口をタンゼンシャルポート２３０ｔとし、燃焼室を半径ＳＲｔの略球面状とし、インジェクタ１２ｔを駆動流通路ではなく、吸気流出通路２３ｔに設けたものである。
インジェクタ１２ｔに供給する酸水素は、酸水素発生手段９ｔの酸水素発生装置９０ｔを制御して酸水素の発生量を調整し、発生した酸水素を流体制御手段６の制御弁６３にて圧力調整し、余剰酸水素は制御弁６３－２にてアキュムレータ６７に適時に貯蔵／放出し、酸水素を安定供給し、回生エネルギを酸水素に変換して貯蔵することにより二次電池の電気容量の負荷を軽減する。
前記酸水素は、インジェクタ１２ｔから吸気流出通路２３ｔに適時に供給し、排気弁４７ｔの閉弁後に燃焼室に流入する吸気に予混合することにより燃料の吹き抜けを防止する。インジェクタ１２ｔ２に供給する火花点火式内燃機関１ｔの主燃料であるガソリン等は、流体供給手段７ｔの燃料タンク７５ｔに貯蔵し、高圧燃料ポンプ１３ｔで加圧して供給し、インジェクタ１２ｔ２にて排気弁４７ｔの閉弁直後に燃焼室に供給し、燃料の吹き抜けを防止する。
燃焼室に供給された前記燃料は、実施例１と同様にスワールの遠心分離作用により水素は点火プラグ１１ｔ付近に集まり、点火プラグ１１ｔにより最小着火エネルギが小さい水素に点火し、燃焼速度が大きい水素のシリンダ軸の中心から周方向に均一な火炎伝播により、主燃料の燃焼を促進して内燃機関１ｔの出力増大効果がある。
シリンダ軸対称の層状燃料層の最外層は極低濃度層となるので、火炎伝播の到達する以前に急激に残りの未燃混合気が燃焼するノッキングが防止でき、前記最外層は燃焼室壁に接する面積が大きいが極低濃度層であるので冷却損失が抑制でき、更に酸水素の酸素は吸気の酸素富化により、熱効率の向上と排気性状の改善の効果がある。
前記酸水素発生手段９ｔは、直流電源を供給する二次電池９６ｔとインバータ９７に並列に接続され、前記インバータ９７にてモータ／発電機９８に電気エネルギの授受を行い、内燃機関１ｔの出力アシストまたはエネルギ回生を行う。
内燃機関１ｔの制御システムは後述する制御システム（図２７）の構成概念の説明図にて、前記ハイブリッド車両のＨＣＣＩエンジンまたはＳＰＣＣＩエンジンへの運転切換え等の内燃機関１ｔの制御方法は後述する制御フローチャート（図２８）にて説明する。 FIG. 27 is an illustration of a concept of a spark-ignition internal combustion engine of a hybrid vehicle provided with a scavenging air amplification means, an oxyhydrogen generator and a regeneration means according to a thirteenth embodiment (corresponding to claim 3).
In the internal combustion engine 1t shown in FIG. 27, the intake port of the internal combustion engine 1s of the eighth embodiment (FIG. 20) is a tangential port 230t, the combustion chamber is substantially spherical with a radius SRt, and the injector 12t is not a drive flow passage. , In the intake and outlet passages 23t.
The acid hydrogen supplied to the injector 12t controls the amount of generated acid hydrogen by controlling the acid hydrogen generator 90t of the acid hydrogen generating means 9t, and the pressure of the generated acid hydrogen is adjusted by the control valve 63 of the fluid control means 6. The excess oxyhydrogen is stored / released to the accumulator 67 at a control valve 63-2 in a timely manner, the oxyhydrogen is stably supplied, the regenerative energy is converted to the oxyhydrogen, and stored. Reduce the load.
The oxyhydrogen is timely supplied from the injector 12t to the intake and outflow passage 23t, and premixed with the intake air flowing into the combustion chamber after the exhaust valve 47t is closed, thereby preventing the blowout of fuel. Gasoline which is the main fuel of the spark ignition type internal combustion engine 1t to be supplied to the injector 12t2 is stored in the fuel tank 75t of the fluid supply means 7t, pressurized by the high pressure fuel pump 13t and supplied, and the injector 12t2 discharges the exhaust valve 47t. Immediately after the valve is closed, it is supplied to the combustion chamber to prevent the blowout of fuel.
In the fuel supplied to the combustion chamber, hydrogen is collected in the vicinity of the spark plug 11t by the centrifugal separation action of the swirl as in the first embodiment, and hydrogen having a small minimum ignition energy is ignited by the spark plug 11t. The uniform flame propagation in the circumferential direction from the center of the cylinder axis promotes the combustion of the main fuel to increase the output of the internal combustion engine 1t.
Since the outermost layer of the cylinder-axisymmetric layered fuel layer is a very low concentration layer, it is possible to prevent knocking that the remaining unburned mixture burns rapidly before reaching the flame propagation, and the outermost layer is on the combustion chamber wall Although the area in contact is large but because it is a very low concentration layer, cooling loss can be suppressed, and oxygen of oxyhydrogen has the effects of improving thermal efficiency and improving exhaust properties by oxygen enrichment of intake air.
The oxy-hydrogen generating unit 9t is connected in parallel to a secondary battery 96t for supplying DC power and an inverter 97, and sends and receives electric energy to the motor / generator 98 by the inverter 97 to assist the output of the internal combustion engine 1t. Or perform energy regeneration.
The control system of the internal combustion engine 1t is an explanatory view of a configuration concept of a control system (FIG. 27) described later, and a control flowchart of the internal combustion engine 1t such as operation switching of the hybrid vehicle to HCCI engine or SPCCI engine will be described later This will be described with reference to FIG.
図２８は、前記実施例１３（図２７）の前記ハイブリッド車両の内燃機関で、ＨＣＣＩエンジンまたはＳＰＣＣＩエンジンとして運転できる制御システムの構成概念の説明図である。
内燃機関１ｔの電子制御装置であるＥＣＵ１９は、ＣＰＵ（中央演算処理装置）、ＲＡＭとＲＯＭからなる記憶素子、入力ポート、出力ポート、およびＤＣ電源で構成され、本図では前記入出力ポートの接続は、中継機器（コントローラ、アンプ、コンバータ等）は図示省略している。
前記ハイブリッド車両の内燃機関１ｔは、クランク角センサ４５、ノックセンサ４８、水温センサ４９、等の入力情報と、前記ハイブリッド車両の制御補助装置であるアクセル開度センサ１７、ブレーキ開度センサ１８、図示しない車速センサ等の入力情報により、ＥＣＵ１９が内燃機関１ｔの運転状況を分析、判断、及び予想して、前記出力手段４ｔの点火プラグ１１ｔとインジェクタ（１２ｔ、１２ｔ２）を内燃機関の運転状況に対応し、掃気増幅手段５ｔの制御弁５６ｔにて掃気量と過給圧を調整し、流体供給手段６の制御弁（６３、６３－２）で酸水素の供給圧を調整し、排気還流通路に設けた制御弁３８ｔにて排気還流量を調整制御する。
前記内燃機関１ｔの制御システムにより、前記ハイブリッド車両の運転状況に対応してＥＣＵ１９が下記制御フローチャート（図２８）に従って内燃機関１ｔの制御を行う。 FIG. 28 is an explanatory view of a configuration concept of a control system which can be operated as an HCCI engine or an SPCCI engine in the internal combustion engine of the hybrid vehicle of the thirteenth embodiment (FIG. 27).
The ECU 19 which is an electronic control unit for the internal combustion engine 1t is composed of a CPU (central processing unit), a storage element consisting of a RAM and a ROM, an input port, an output port, and a DC power source. The relay device (controller, amplifier, converter, etc.) is not shown.
The internal combustion engine 1t of the hybrid vehicle includes input information such as a crank angle sensor 45, a knock sensor 48, a water temperature sensor 49, etc., an accelerator opening sensor 17 which is a control assisting device of the hybrid vehicle, a brake opening sensor 18, The ECU 19 analyzes, judges, and predicts the operating condition of the internal combustion engine 1t based on input information from the vehicle speed sensor etc., and adapts the spark plug 11t and the injector (12t, 12t2) of the output means 4t to the operating condition of the internal combustion engine The control valve 56t of the scavenging amplification means 5t adjusts the scavenging amount and the supercharging pressure, and the control valve (63, 63-2) of the fluid supply means 6 adjusts the supply pressure of oxyhydrogen to the exhaust gas recirculation passage. The exhaust gas recirculation amount is adjusted and controlled by the control valve 38t provided.
The control system of the internal combustion engine 1t causes the ECU 19 to control the internal combustion engine 1t according to the following control flowchart (FIG. 28) in accordance with the driving condition of the hybrid vehicle.
図２９は、前記実施例１３の内燃機関１ｔ（図２７）を組み込んだ前記ハイブリッド車両の内燃機関の制御システム（図２８）を、予混合圧縮自動着火のＨＣＣＩエンジンまたは火花制御による圧縮着火燃焼のＳＰＣＣＩエンジンとして運転する制御フローチャート（駆動アシスト等のモータ制御は除く）である。
図２９は、前記制御システム（図２８）の入出力情報を演算処理するＥＣＵ１９により制御される。
加速あるいは減速等の判断は、主にアクセルあるいはブレーキのペダル操作によるアクセル開度センサ１７、ブレーキ開度センサ１８、および図示しない車速センサ等からの入力情報により、運転者の意思や内燃機関１ｔの運転状況を分析、判断、予測し、各運転サブルーチンを本制御フローチャートに従い実行し、出力ポートからの出力によりアクチェータ等を制御する。
前記制御フローチャートは、内燃機関１ｔの制御の説明であるので、エネルギ回生の駆動アシスト等のモータ制御や酸水素発生手段の電気制御関連等は説明を省略する。
まず、ＥＣＵ１９は、アクセル開度センサ１７、ブレーキ開度センサ１８、前記車速センサ等により、加速のための燃焼運転が必要であるかを判断する（ステップＳ１０）。
ここで、燃焼運転が必要であると判断した場合は、内燃機関１ｔの過給圧、水温等よりＴＤＣでの燃焼室の吸気温度の予測と、ノックセンサ４８の情報等によりＨＣＣＩエンジンまたはＳＰＣＣＩエンジンとして運転が可能かを判断する（ステップＳ１１）。
一方、燃焼運転が必要でないと判断した場合は、ブレーキ開度センサ１７がＯＮであるかを判断する（ステップＳ１２）。
ここで、ブレーキ開度センサ１７がＯＮであると判断した場合は、エネルギ回生が可能かを判断する（ステップＳ１３）。
具体的には、前記車速センサ等により回生できる運動エネルギと予測される減速量によりエネルギ回生の可否判断を行う。一方、ブレーキ開度センサ１７がＯＮでないと判断した場合は、積極的に加速も減速もしない慣性運転（フリーラン）サブルーチン（ステップＳ１４）を実行した後、ＲＥＴＵＲＮにて本処理ルーチンを一旦終了する。
前記エネルギ回生が可能かの判断（ステップＳ１３）で、エネルギ回生が可能であると判断した場合は、エネルギ回生サブルーチン（Ｓ１５）を実行し、インバータ９７にてモータ／発電機９８で発電される電力を酸水素発生装置９０ｔと二次電池９６ｔに供給してエネルギ回生を行い、モータ／発電機９８の逆起電力によるトルクを制動に利用する。
一方、エネルギ回生が可能でないと判断した場合は、エンジンブレーキサブルーチン（ステップＳ１６）を実行する。
具体的には、燃料供給を停止し、圧縮仕事やポンピングロス等が発生するように掃気増幅手段５ｔの制御弁５６を制御する。 FIG. 29 is a control system (FIG. 28) of an internal combustion engine of the hybrid vehicle incorporating the internal combustion engine 1t (FIG. 27) of the thirteenth embodiment, in compression ignition combustion by HCCI engine or spark control of homogeneous charge compression auto ignition. It is a control flow chart (except motor control, such as drive assist) which operates as a SPCCI engine.
FIG. 29 is controlled by the ECU 19 that processes input / output information of the control system (FIG. 28).
The determination of acceleration or deceleration is made mainly by the driver's intention or the internal combustion engine 1t by input information from the accelerator opening sensor 17, brake opening sensor 18 and vehicle speed sensor (not shown) by accelerator or brake pedal operation. The operation situation is analyzed, judged and predicted, and each operation subroutine is executed according to this control flowchart, and an actuator etc. is controlled by the output from the output port.
Since the control flow chart is a description of control of the internal combustion engine 1t, description of motor control such as drive assist for energy regeneration and electric control related to oxyhydrogen generation means is omitted.
First, the ECU 19 determines whether a combustion operation for acceleration is required by the accelerator opening sensor 17, the brake opening sensor 18, the vehicle speed sensor, etc. (step S10).
Here, when it is determined that the combustion operation is necessary, the HCCI engine or the SPCCI engine is predicted from the supercharge pressure of the internal combustion engine 1t, the water temperature of the internal combustion engine, etc. It is determined whether the operation is possible (step S11).
On the other hand, when it is determined that the combustion operation is not necessary, it is determined whether the brake opening degree sensor 17 is ON (step S12).
Here, if it is determined that the brake opening sensor 17 is ON, it is determined whether energy regeneration is possible (step S13).
Specifically, whether or not energy regeneration is possible is determined based on the kinetic energy that can be regenerated by the vehicle speed sensor or the like and the estimated amount of deceleration. On the other hand, if it is determined that the brake opening sensor 17 is not ON, an inertial operation (free run) subroutine not actively accelerating or decelerating is executed (step S14), and this processing routine is temporarily ended at RETURN. .
If it is determined that energy regeneration is possible in the determination of whether the energy regeneration is possible (step S13), the energy regeneration subroutine (S15) is executed, and the electric power generated by the motor / generator 98 by the inverter 97 Is supplied to the oxyhydrogen generator 90t and the secondary battery 96t to perform energy regeneration, and the torque by the back electromotive force of the motor / generator 98 is used for braking.
On the other hand, when it is determined that energy regeneration is not possible, an engine brake subroutine (step S16) is executed.
Specifically, the fuel supply is stopped, and the control valve 56 of the scavenging amplification means 5t is controlled so that compression work, pumping loss and the like occur.
前記（ステップＳ１１）で、前記ＴＤＣでの燃焼室の吸気の圧力と温度予測等を基にＨＣＣＩ運転またはＳＰＣＣＩ運転が可能でないと判断した場合は、掃気増幅手段５ｔの制御弁５６ｔ、排気還流通路３６の制御弁３８等を制御して内燃機関１ｔの燃焼室等の運転状況をＨＣＣＩ運転またはＳＰＣＣＩ運転が可能なるように機関運転調整サブルーチン（ステップ１７）を実行し、主燃料が必要かを判断する（ステップ２１）。
前記（ステップ１１）で内燃機関１ｔのＨＣＣＩ運転またはＳＰＣＣＩ運転が可能であると判断した場合は、水素燃料が供給可能かを判断する（ステップ１８）。
ここで、水素燃料が供給可能でないと判断した場合は、内燃機関１ｔの主燃料によるＨＣＣＩまたはＳＰＣＣＩ運転サブルーチン（ステップ１９）を実行した後、ＲＥＴＵＲＮにて本処理ルーチンを一旦終了する。
具体的には、掃気増幅手段５ｔの制御弁５６ｔ、排気還流通路３６の制御弁３８等を制御して内燃機関１ｔのＴＤＣでの吸気温度を主燃料の発火点以上にする、または前記吸気温度を主燃料の発火点附近にし、点火プラグ１１ｔで点火するように運転調整する。
一方、前記（ステップ１８）で水素燃料が供給可能であると判断した場合は、内燃機関１ｔの水素燃料、または主燃料と水素燃料の二燃料によるＨＣＣＩまたはＳＰＣＣＩ運転サブルーチン（ステップ２０）を実行した後、ＲＥＴＵＲＮにて本処理ルーチンを一旦終了する。
前記主燃料が必要かを判断する（ステップ２１）で、主燃料が必要であると判断した場合は、水素燃料が供給可能かを判断する（ステップ２２）。
一方、水素燃料が供給可能であると判断した場合は、内燃機関１ｔの主燃料と水素燃料の二燃料による火花点火式内燃機関運転サブルーチン（ステップ２４）を実行した後、ＲＥＴＵＲＮにて本処理ルーチンを一旦終了する。
ここで、水素燃料が供給可能でないと判断した場合は、内燃機関１ｔの水素燃料による火花点火式内燃機関運転サブルーチン（ステップ２５）を実行した後、ＲＥＴＵＲＮにて本処理ルーチンを一旦終了する。
以上の制御フローに従って、ハイブリッド車両の火花点火式内燃機関１ｔの運転状況に応じて、前記内燃機関１ｔの燃料供給、過給、ＥＧＲ、エネルギ回生等を各サブルーチンに従って実行する。
以上のようにＥＣＵ１９の入出力情報により、前記ハイブリッド車両の内燃機関１ｔの制御システム（図２４）を制御し、本制御フローチャートは、前記内燃機関１ｔの運転中は繰り返し実行される。 If it is determined in step S11 that HCCI operation or SPCCI operation is not possible based on the pressure and temperature prediction of the intake air of the combustion chamber at TDC, the control valve 56t of the scavenging gas amplification means 5t, the exhaust gas recirculation passage Execute the engine operation adjustment subroutine (step 17) so that HCCI operation or SPCCI operation is possible by controlling the control valve 38 etc. of 36 and operating the combustion chamber etc. of the internal combustion engine 1t, and determine whether the main fuel is necessary. (Step 21).
If it is determined in the step 11 that the HCCI operation or the SPCCI operation of the internal combustion engine 1t is possible, it is determined whether the hydrogen fuel can be supplied (step 18).
Here, if it is determined that the hydrogen fuel can not be supplied, the HCCI or SPCCI operation subroutine (step 19) by the main fuel of the internal combustion engine 1t is executed, and the processing routine is temporarily ended at RETURN.
Specifically, the control valve 56t of the scavenging gas amplification means 5t, the control valve 38 of the exhaust gas recirculation passage 36, and the like are controlled to make the intake temperature at TDC of the internal combustion engine 1t equal to or higher than the ignition point of the main fuel, or the intake temperature And adjust the operation so as to ignite with the spark plug 11t near the ignition point of the main fuel.
On the other hand, if it is determined that hydrogen fuel can be supplied in the above (step 18), the HCCI or SPCCI operation subroutine (step 20) is executed using the hydrogen fuel of the internal combustion engine 1t or two fuels of the main fuel and hydrogen fuel. After that, this processing routine is temporarily ended at RETURN.
If it is judged at step 21 that the main fuel is necessary, then it is judged whether hydrogen fuel can be supplied if it is judged that the main fuel is necessary (step 22).
On the other hand, if it is determined that hydrogen fuel can be supplied, a spark ignition type internal combustion engine operation subroutine (step 24) is executed by the two fuels of the main fuel of the internal combustion engine 1t and hydrogen fuel, and then the processing routine is executed at RETURN. Once.
Here, if it is determined that the hydrogen fuel can not be supplied, a spark ignition type internal combustion engine operation subroutine (step 25) using the hydrogen fuel of the internal combustion engine 1t is executed, and this processing routine is temporarily ended at RETURN.
According to the control flow described above, fuel supply, supercharging, EGR, energy regeneration and the like of the internal combustion engine 1t are executed according to each subroutine according to the operating condition of the spark ignition internal combustion engine 1t of the hybrid vehicle.
As described above, the control system (FIG. 24) of the internal combustion engine 1t of the hybrid vehicle is controlled by the input / output information of the ECU 19, and this control flowchart is repeatedly executed during operation of the internal combustion engine 1t.
図３０は、実施例１４（請求項４対応）の、略球形の燃焼室に吸気弁と排気弁を放射状に配置し、２本のカム軸の各カムで各排気弁の開閉と油圧手段を介して各吸気弁を開閉する４サイクル内燃機関の構成概念の説明図で、上図は内燃機関１ａの平面図で、下図は上図のＡ－Ａ断面の部分断面を含む内燃機関１ａの断面図である。
図３０は、４サイクル内燃機関１ａにおいて、燃焼室を半径ＳＲａの略球面状とし、前記燃焼室に放射状に複数の吸気弁４６ａと排気弁４７ａを交互に配置し、インジェクタ１２ｄを前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポート２３０ａとし、水素のように空気より密度が小さい燃料を前記燃焼室に供給し、内燃機関１ｄの運転状況に応じて前記燃料の供給を制御し、更に、クランク軸４４ａの回転数の１／２の回転数で連動する平行な２本のカム軸４０７（ａ１，ａ２）を設け、前記排気弁４７ａを前記２本の各カム軸４０７（ａ１，ａ２）に設けたカム４０８（ａ１，ａ２）により開閉し、前記吸気弁を前記カム４０８（ａ１，ａ２）に連動する油圧手段であるプランジャ８４ａ（ａ１，ａ２）、弁シリンダ４７１（ａ１，ａ２）、弁ピストン４７２（ａ１，ａ２）により開閉する４サイクル内燃機関１ａである。
内燃機関１ａは、クランク軸４４ａに設けた駆動車４０１ａで伝動媒体４０３ａを介してカム軸４０７ａ１に設けた駆動車４０１ａの有効径φＤａの２倍の有効径（φ２Ｄａ）の従動車４０２ａを駆動し、カム軸４０７ａ１をクランク軸４４ａの２分の１の回転数で回転する。
前記カム軸４０７ａ１に設けた駆動歯車４０５ａに噛合うカム軸４０７ａ２に設けた前記駆動歯車４０５ａとピッチ円直径（φｄａ）が同じ従動歯車４０６ａにより、カム４０８（ａ１、ａ３）を設けたカム軸４０７ａ１とカム４０８（ａ２、ａ４）を設けたカム軸４０７ａ２は同一回転数で逆方向に回転する。
カム４０８ａ１とカム４０８ａ２は、シリンダ軸に対し左右対称のカム形状とし、それぞれのカムで作動する各排気弁４７ａはクランク軸４４ａに同期して開弁する。 In FIG. 30, the intake valve and the exhaust valve are radially arranged in the substantially spherical combustion chamber of the fourteenth embodiment (corresponding to claim 4), and the opening and closing of each exhaust valve and the hydraulic means are realized by the respective cams of two camshafts. FIG. 5 is an explanatory view of a construction concept of a four-stroke internal combustion engine for opening and closing each intake valve, the upper view is a plan view of the internal combustion engine 1a, and the lower view is a cross section of the internal combustion engine 1a FIG.
In FIG. 30, in the four-stroke internal combustion engine 1a, the combustion chamber has a substantially spherical shape with a radius SRa, a plurality of intake valves 46a and exhaust valves 47a are arranged alternately in the combustion chamber, and the injector 12d is of the combustion chamber. The intake port is provided near the intersection with the cylinder axis, and the intake port is a tangential port 230a that generates swirl in the cylinder, and fuel such as hydrogen having a density smaller than that of air is supplied to the combustion chamber to operate the internal combustion engine 1d In accordance with the above, the supply of the fuel is controlled, and further, two parallel camshafts 407 (a1, a2) interlocking at a rotational speed of 1/2 of the rotational speed of the crankshaft 44a are provided, and the exhaust valve 47a is A pump which is a hydraulic means that opens and closes by means of cams 408 (a1, a2) provided on the two respective camshafts 407 (a1, a2) and interlocks the intake valve with the cams 408 (a1, a2) Nja 84a (a1, a2), the valve cylinder 471 (a1, a2), is a 4-cycle internal combustion engine 1a for opening and closing the valve piston 472 (a1, a2).
The internal combustion engine 1a drives a driven vehicle 402a having an effective diameter (.phi.2Da) twice the effective diameter .phi.Da of the drive vehicle 401a provided on the camshaft 407a1 via the transmission medium 403a by the drive vehicle 401a provided on the crankshaft 44a. The cam shaft 407a1 is rotated at a half rotation speed of the crank shaft 44a.
A cam shaft 407a1 provided with a cam 408 (a1, a3) by a driven gear 406a having the same pitch circle diameter (φda) as the drive gear 405a provided on the cam shaft 407a2 engaged with the drive gear 405a provided on the cam shaft 407a1. The cam shaft 407a2 provided with the cams 408 (a2, a4) rotates in the opposite direction at the same rotational speed.
The cams 408a1 and the cams 408a2 have cam shapes symmetrical with respect to the cylinder axis, and the exhaust valves 47a operated by the respective cams are opened in synchronization with the crankshaft 44a.
吸気弁４６ａは、前記カム４０８（ａ１，ａ２）に設けた排気弁４７ａより位相が遅れて作動するプランジャ８４（ａ１、ａ２）で発生する油圧で、対応する弁シリンダ４７１（ａ１，ａ２）の弁ピストン４７２（ａ１，ａ２）を作動して弁の開閉動作を行う。
燃焼室に放射状に複数の吸気弁４６ａと排気弁４７ａを交互に配置し、更に吸気弁４６ａ同士と排気弁４７ａ同士は同一線上に配置し、下図に示すように吸気弁４６ａと排気弁４７ａの狭角をθａ（θａ>９０ー）とすることにより、多気筒の弁の配置干渉を軽減し、気筒間の距離を短縮することができ、小型軽量で、剛性の大きいシリンダブロックにできる。
内燃機関１ａのタンゼンシャルポート２３０ａとインジェクタ１２ａから供給される水素のように空気より密度が小さい燃料の遠心分離作用は前記２サイクル内燃機関の実施例１～３と基本原理は同じであるので説明を省略する。
排気弁４７ａは排気行程の排気圧が働く状態での開弁には、排気圧による推力と弁スプリングによる付勢力に抗する弁推力が必要であり開弁には大きな力が必要であるので、カム機構のように剛性が大きく、高速運転でのキャビテーション等の問題が発生しない機械式弁駆動機構が適している。
吸気弁４６ａは、排気弁４７ａにより略大気圧となった排気残圧とスプリング４７３ａに抗する弁推力が必要であるが、排気弁の弁推力と比較して小さな弁推力でよい。 The intake valve 46a is a hydraulic pressure generated by a plunger 84 (a1, a2) that operates with a phase lag behind the exhaust valve 47a provided to the cam 408 (a1, a2). The valve piston 472 (a1, a2) is operated to open and close the valve.
A plurality of intake valves 46a and exhaust valves 47a are alternately arranged radially in the combustion chamber, and the intake valves 46a and the exhaust valves 47a are arranged on the same line, and as shown in the figure below, the intake valves 46a and the exhaust valves 47a are By setting the narrow angle to θa (θa> 90−), the arrangement interference of multi-cylinder valves can be reduced, the distance between cylinders can be shortened, and a compact, lightweight, rigid cylinder block can be obtained.
The centrifugal separation action of fuel having a density smaller than air, such as hydrogen supplied from the tangential port 230a of the internal combustion engine 1a and the injector 12a, has the same basic principle as the first to third embodiments of the two-stroke internal combustion engine. I omit explanation.
Since the exhaust valve 47a requires a thrust by the exhaust pressure and a valve thrust against the biasing force by the valve spring to open the valve in a state where the exhaust pressure in the exhaust stroke works, a large force is necessary to open the valve. A mechanical valve drive mechanism that has high rigidity and does not cause problems such as cavitation at high speed operation like a cam mechanism is suitable.
The intake valve 46a needs an exhaust residual pressure that is substantially atmospheric pressure by the exhaust valve 47a and a valve thrust that resists the spring 473a, but may have a valve thrust that is smaller than the valve thrust of the exhaust valve.
図３１は、前記実施例１４（図３０）の、ＴＤＣの水素の各燃料濃度層の分布状況と燃料噴射時のＨＣＣＩエンジンの燃焼の説明図である。
図３１は、前記実施例１４（図３０）の前記空気より密度が小さい燃料の遠心分離作用により、インジェクション１２ａより供給される予混合燃料である水素のＴＤＣでの燃料濃度層の分布状況である。
ＴＤＣにて断熱圧縮されて軽油の発火点（２５０℃）以上の温度の水素予混合気に、燃焼室とシリンダ軸との交点近傍に設けたインジェクション１２ａから噴射された主燃料である軽油は、水素の高濃度層（Ｆ１ａ）にて霧化と気化をしながら自己着火してシリンダ軸付近の水素の高濃度層（Ｆ１ａ）が高速燃焼を開始するので、スワールにより高速回転しているシリンダ軸中心から高温高圧の火炎伝播が周方向に均一に膨張し、更に中濃度層（Ｆ２ａ）の水素も軽油の燃焼と前記火炎伝播により燃焼を開始するので燃焼室内の圧力と温度の上昇により水素の発火点（５８５℃）を超え、低濃度層（Ｆ３ａ）等の水素燃焼が拡散している領域の水素も着火することにより燃焼室全域の燃焼が促進されてパティキュレイト、デポジット等の発生を抑制する。
前記周方向に均一な燃焼により未燃焼ガスの発生が少なく、外周層である予混合水素の超低濃度層（Ｆ４ａ）でのノッキングが抑制され、燃焼室壁面との接触面積が大きい前記外周層（Ｆ４ａ）の発熱量が小さいので冷却損失が小さいので内燃機関１ａの熱効率が向上する効果がある。 FIG. 31 is an explanatory view of a distribution state of fuel concentration layers of hydrogen of TDC and combustion of an HCCI engine at the time of fuel injection in the fourteenth embodiment (FIG. 30).
FIG. 31 is a distribution state of a fuel concentration layer at TDC of hydrogen, which is a premixed fuel supplied from the injection 12a, by the centrifugal separation action of the fuel having a density smaller than that of the air in the fourteenth embodiment (FIG. 30). .
The diesel fuel, which is the main fuel injected from the injection 12 a near the intersection of the combustion chamber and the cylinder axis, is a hydrogen premixed gas that is adiabatically compressed at TDC and at a temperature above the ignition point (250 ° C.) of the diesel. Since the high concentration layer of hydrogen (F1a) near the cylinder axis starts high-speed combustion while self-igniting while atomizing and vaporizing in the high concentration layer of hydrogen (F1a), the cylinder shaft rotating at high speed by swirling The high temperature and high pressure flame propagation from the center expands uniformly in the circumferential direction, and hydrogen in the middle concentration layer (F2a) also starts combustion by the combustion of the light oil and the flame propagation, so the pressure and temperature of the hydrogen in the combustion chamber increase. Combustion in the entire combustion chamber is promoted by igniting hydrogen in a region where hydrogen combustion is diffused, such as low concentration layer (F3a), exceeding the ignition point (585 ° C), and generation of particulates, deposits, etc. Suppress.
The outer circumferential layer has less generation of unburned gas due to the uniform combustion in the circumferential direction, suppresses knocking in the ultra-low concentration layer (F4a) of premixed hydrogen as the outer circumferential layer, and has a large contact area with the wall surface of the combustion chamber. Since the amount of heat generation of (F4a) is small and the cooling loss is small, there is an effect of improving the thermal efficiency of the internal combustion engine 1a.
図３２は、実施例１５（請求項４対応）の、２本のカム軸に設けた各カムにより排気弁を開閉し、前記カムとは別のカムにより油圧手段を介して吸気弁を開閉する４サイクル内燃機関の平面図と周辺回路図である。
図３２は、４サイクル内燃機関１ｂにおいて、燃焼室を半径ＳＲｂの略球面状とし、前記燃焼室に放射状に複数の吸気弁４６（ｂ１～ｂ４）と排気弁４７（ｂ１～ｂ４）を交互に配置し、点火プラグ１１（ｂ１、ｂ２）を前記燃焼室のシリンダ軸との交点近傍に設け、吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポート２３０（ｂ１～ｂ４）とし、水素のように空気より密度が小さい燃料と火花点火式内燃機関の燃料であるガソリンを吸気系統および前記燃焼室に供給し、前記内燃機関の運転状況に応じて前記燃料の供給を制御し、更に、クランク軸４４ｂの回転数の１／２の回転数で連動する平行な２本のカム軸４０７（ｂ１、ｂ２）を設け、前記排気弁４７（ｂ１～ｂ４）を前記２本の各カム軸４０７（ｂ１、ｂ２）に設けたカム４０８（ｂ１、ｂ２）により開閉し、前記吸気弁４６（ｂ１～ｂ４）を前記カム４０８ｂ１に連動する油圧手段である弁シリンダ４７１ｂ２により開閉するまたは前記吸気弁４６（ｂ１～ｂ４）を前記カム４０８とは別のカムに連動する油圧手段である弁駆動ユニット８０（ｂ１、ｂ３）により開閉する４サイクル内燃機関１ｂである。
クランク軸４４ｂに設けた駆動車４０１ｂと伝動媒体４０３ｂを介してカム軸４０７ｂ１に設けた従動車４０２ｂへの１/２に減速する回転駆動、およびカム軸４０７ｂ１に設けた駆動歯車４０５ｂと噛合うカム軸４０７ｂ２に設けた従動歯車４０６ｂによる等速回転駆動の方法は実施例１４と同様であるので説明を省略する。
内燃機関１ｂのタンゼンシャルポート２３０ａとインジェクタ１２ａから供給される水素のように空気より密度が小さい燃料の遠心分離作用は前記２サイクル内燃機関の実施例１～３と同じであるので説明を省略する。 32. In FIG. 32, the exhaust valve is opened and closed by the cams provided on the two camshafts of the fifteenth embodiment (corresponding to claim 4), and the intake valve is opened and closed through the hydraulic means by the cam other than the cam. FIG. 1 is a plan view and a peripheral circuit diagram of a four-stroke internal combustion engine.
In FIG. 32, in the four-stroke internal combustion engine 1b, the combustion chamber has a substantially spherical shape with a radius SRb, and a plurality of intake valves 46 (b1 to b4) and exhaust valves 47 (b1 to b4) are alternately arranged radially to the combustion chamber. Place the spark plug 11 (b1, b2) in the vicinity of the intersection with the cylinder axis of the combustion chamber, and use the intake port as the tangential port 230 (b1 to b4) that generates swirl in the cylinder. Supplying a fuel having a density smaller than that of air and a fuel of a spark ignition type internal combustion engine to an intake system and the combustion chamber, and controlling the supply of the fuel according to the operating condition of the internal combustion engine; Two parallel camshafts 407 (b1, b2) interlocking at a rotational speed of 1⁄2 of 44b are provided, and the exhaust valves 47 (b1 to b4) are connected to the respective camshafts 407 (b1) , B2) The intake valve 46 (b1 to b4) is opened or closed by a valve cylinder 471b2, which is a hydraulic means interlocked with the cam 408b1 and opened or closed by the cam 408 (b1, b2) provided. The four-stroke internal combustion engine 1b is opened and closed by a valve drive unit 80 (b1, b3) which is hydraulic means interlocking with a cam different from the cam 408.
A rotational drive that decelerates to a half of the drive wheel 401b provided on the crankshaft 44b and the driven vehicle 402b provided on the camshaft 407b1 via the transmission medium 403b, and a cam that meshes with the drive gear 405b provided on the camshaft 407b1 The method of constant-velocity rotational drive by the driven gear 406b provided on the shaft 407b2 is the same as that of the fourteenth embodiment, and thus the description thereof is omitted.
The centrifugal separation action of fuel having a density smaller than air, such as hydrogen supplied from the tangential port 230a of the internal combustion engine 1b and the injector 12a, is the same as in the first to third embodiments of the two-stroke internal combustion engine Do.
図３３は、前記実施例１５（図３２）の、ＴＤＣの各燃料濃度層の分布状況と水素可燃層の点火時のＳＰＣＣＩエンジン（火花制御による圧縮着火燃焼）の説明図である。
ＴＤＣにて断熱圧縮されてガソリンの発火点（３００℃）以下の温度の水素予混合気に、燃焼室とシリンダ軸との交点近傍に設けた点火プラグ１１ｂ１で火花点火し、シリンダ軸中心付近の水素高濃度層（Ｆ１ｂ）が高速燃焼を開始するので、スワールにより高速回転しているシリンダ軸中心から高温高圧の火炎伝播が周方向に均一に膨張し、更に中濃度層（Ｆ２ｂ）に拡散している水素とガソリンも前記火炎伝播により燃焼を開始するので燃焼室内の圧力と温度の上昇により水素の発火点（５８５℃）を超え、低濃度層（Ｆ３ｂ）等の水素とガソリンが拡散している領域の水素も着火することにより燃焼室全域の燃焼が促進されて未燃焼ガスの発生を抑制する。
外周層である予混合水素の超低濃度層Ｆ４ｂでのノッキング抑制と内燃機関１ｂの熱効率向上効果の説明は、実施例１４（図３１）と同じであるので説明を省略する。 FIG. 33 is an explanatory view of a distribution state of fuel concentration layers of TDC and an SPCCI engine (compression ignition combustion by spark control) at the time of ignition of a hydrogen combustible layer in the fifteenth embodiment (FIG. 32).
A hydrogen premixed gas which is adiabatically compressed at TDC and has a temperature below the ignition point (300 ° C) of gasoline is spark-ignited by the spark plug 11b1 provided near the intersection of the combustion chamber and the cylinder axis. Since the high hydrogen concentration layer (F1b) starts high-speed combustion, flame propagation at high temperature and high pressure expands uniformly in the circumferential direction from the center of the cylinder axis rotating at high speed by swirl and diffuses further to the middle concentration layer (F2b) Since hydrogen and gasoline also start combustion by the flame propagation, the pressure and temperature in the combustion chamber increase the hydrogen ignition point (585 ° C) and the hydrogen and gasoline such as low concentration layer (F3b) diffuses. Ignition of hydrogen in the region where it is located promotes combustion in the entire combustion chamber to suppress the generation of unburned gas.
The explanation of the knocking suppression in the ultra-low concentration layer F4b of the premixed hydrogen which is the outer peripheral layer and the thermal efficiency improvement effect of the internal combustion engine 1b is the same as that of the fourteenth embodiment (FIG. 31), so the description is omitted.
前記実施例１～１５は、本願発明の一例を説明したもので、各実施例の内燃機関は、制約のない限り、ディーゼル機関でも火花点火式内燃機関でもよく、燃料の供給は制約のない限り吸気系統でも燃焼室でもよく、過給増幅手段の空気流量増幅器は、エジェクタ、フロートランスベクタ、トランスベクタ等のいずれであってもよく、圧縮機は往復圧縮機で説明したが、リショルム・コンプレッサ等の他の圧縮機でもよく、ハイブリッド車両はパラレルでもシリーズでもよい。
前記実施例１～１５は、本願発明の一例を示すもので本願発明を制約するものではなく、当業者により変更および改良ができる。 The embodiments 1 to 15 describe an example of the present invention, and the internal combustion engine of each embodiment may be a diesel engine or a spark ignition internal combustion engine unless restricted, and the supply of fuel is not restricted. The air flow amplifier may be an intake system or a combustion chamber, and the air flow amplifier of the supercharging amplification means may be any of an ejector, a flow transformer vector, a transformer vector, etc. The compressor has been described as a reciprocating compressor. The hybrid vehicles may be parallel or series.
The embodiments 1 to 15 show an example of the present invention and do not limit the present invention, and those skilled in the art can change and improve the present invention.
本発明の内燃機関は、潤滑油の混合を必要としない２サイクル内燃機関で、４サイクル内燃機関と同等のメンテナンス性を有し、簡素な構成で確実で良好な燃焼ができる。
簡素な掃気増幅手段により小さな容量の往復圧縮機で十分な掃気と過給を行えるので、掃気の改善により燃焼性が向上して排気性状が改善し、内燃機関の装置容積当たりの出力が増大して内燃機関をダウンサイジング（小型、軽量化）できるので、自動車、船舶等の移動体に搭載する内燃機関に利用できる。請求項４の４サイクル内燃機関も、内燃機関の燃焼効率の向上により、内燃機関をダウンサイジングして小型、軽量化できる。 The internal combustion engine according to the present invention is a two-stroke internal combustion engine that does not require mixing of lubricating oil, has the same maintainability as a four-stroke internal combustion engine, and can perform reliable and satisfactory combustion with a simple configuration.
Since sufficient scavenging and supercharging can be performed with a small displacement reciprocating compressor by simple scavenging amplification means, the improvement of scavenging improves the combustibility and the exhaust characteristics, and the output per unit volume of the internal combustion engine increases. Since the internal combustion engine can be downsized (small and lightweight), it can be used for an internal combustion engine mounted on a moving body such as a car or a ship. Also in the four-stroke internal combustion engine of claim 4, the internal combustion engine can be downsized and reduced in size and weight by the improvement of the combustion efficiency of the internal combustion engine.
１　内燃機関
２（吸気系統）
３（排気系統）
４　出力手段
５　掃気増幅手段
６　流体制御手段
７　流体供給手段
８　容積型油圧供給手段
９　酸水素発生手段
１１　点火プラグ
１２　インジェクタ
１３　高圧燃料ポンプ
１４　フュエルレール
１５　逆止弁
１６　クランク角センサ
１７　アクセル開度センサ
１８　ブレーキ開度センサ
１９　ＥＣＵ（エンジンコントロールユニット）
２０　吸気
２１　エアクリーナ
２２　吸気流入通路
２３　吸気流出通路
２４　吸気副通路
２５　往復圧縮機
２７　制御弁
２８　過給センサ
２９　アキュムレータ
３１　排気通路
３２　排気浄化装置
３３　消音器
３４　排気センサ
３６　排気還流通路
３７　逆止弁
３８　制御弁
４０　弁機構
４１　シリンダ
４２　ピストン
４３　連結棒
４４　クランク軸
４５　クランク角センサ
４６　吸気弁
４７　排気弁
４８　ノックセンサ
４９　水温センサ
５０　空気流量増幅器
５１　エジェクタ
５２　フロートランスベクタ
５３　トランスベクタ
５５　逆止弁
５６　制御弁
５８　駆動流通路
６１　燃料通路
６２　燃料センサ
６３　制御弁
６４　減圧弁
６７　アキュムレータ
６８　逆止弁
７１　燃料通路
７２　緊急遮断弁
７５　燃料タンク
７６　逆止弁付継手
７７　充填口
８０　弁駆動ユニット（容積型油圧式）
８１　カム
８２　ロータ
８３　ベーン
８４　プランジャ
８５　回転伝動手段
８６　回転継手
８７　油圧補助手段
８８　油圧通路（弁駆動）
８９　油圧通路
９０　酸水素発生装置
９１　電気分解手段
９２　超音波発振手段
９４　電解液タンク
９５　燃料通路
９６　二次電池
９７　インバータ
９８　モータ／発電機　
１３１　サプライポンプ
１４１　コモンレール
２００　過給吸気
２３０　タンゼンシャルポート
２５０　駆動流
２５１　シリンダ（圧縮機）
２５２　ピストン（圧縮機）
２５３　連結棒（圧縮機）
２５４　クランク軸（圧縮機）
２５６　吸入弁
２５７　吐出弁
２５８　駆動歯車
２５９　従動歯車
４０１　駆動車
４０２　従動車
４０３　伝動媒体
４０５　駆動歯車
４０６　従動歯車
４０７　カム軸
４０８　カム
４１０　燃焼室
４２０　球面状キャビティ
４２１　バルブリセス
４７１　弁シリンダ
４７２　弁ピストン
４７３　スプリング
５３１　ノズル
５３２　環状チャンバ
５３３　ハウジング　
５３４　ピストン
５３５　スプリング
５３６　フランジ
５５１　リードバルブ
５５２　リード
５５５　リフト逆止弁
５５６　スプリング
５５７　ディスク
５８１　冷却器
８１０　基準プロフィール
８１１　カムプロフィール
８２１　油圧中継路（第１油圧）
８２１　油圧中継路（第２油圧）
８２１　油圧中継路（第３油圧）
８２６　油圧中継路
８５１　駆動車
８５２　従動車
８５３　伝動媒体　
８７１　油タンク
８７５　逆止弁
９１１　陰極
９１２　陽極
９１３　直流電源
９１４　電解槽
９１５　電解液
９１６　センサ（液面、圧力）
９１７　ポンプ
９１８　制御弁（酸水素）
９２１　超音波発振子
９２２　高周波発生器 1 Internal combustion engine 2 (intake system)
3 (exhaust system)
DESCRIPTION OF SYMBOLS 4 Output means 5 Scavenging amplification means 6 Fluid control means 7 Fluid supply means 8 Volume-type hydraulic pressure supply means 9 Acid hydrogen generation means 11 Spark plug 12 Injector 13 High-pressure fuel pump 14 Fuel rail 15 Check valve 16 Crank angle sensor 17 Accelerator opening degree Sensor 18 Brake opening sensor 19 ECU (engine control unit)
Reference Signs List 20 intake air 21 air cleaner 22 intake inflow passage 23 intake outflow passage 24 intake auxiliary passage 25 reciprocating compressor 27 control valve 28 supercharging sensor 29 accumulator 31 exhaust passage 32 exhaust purification device 33 silencer 34 exhaust sensor 36 exhaust return passage 37 check valve 38 Control valve 40 Valve mechanism 41 Cylinder 42 Piston 43 Connecting rod 44 Crankshaft 45 Crank angle sensor 46 Intake valve 47 Exhaust valve 48 Knock sensor 49 Water temperature sensor 50 Air flow amplifier 51 Ejector 52 Flow transformer vector 53 Transvector 55 Check valve 56 Control valve 58 Drive flow passage 61 Fuel passage 62 Fuel sensor 63 Control valve 64 Pressure reducing valve 67 Accumulator 68 Check valve 71 Fuel passage 72 Emergency shut off valve 75 Fuel tank 76 Joint with check valve 77 Filling port 80 Valve drive unit (volume type Hydraulic type)
81 cam 82 rotor 83 vane 84 plunger 85 rotational transmission means 86 rotary joint 87 hydraulic auxiliary means 88 hydraulic passage (valve drive)
89 hydraulic passage 90 oxyhydrogen generator 91 electrolysis means 92 ultrasonic oscillation means 94 electrolyte solution tank 95 fuel passage 96 secondary battery 97 inverter 98 motor / generator
131 supply pump 141 common rail 200 supercharged intake 230 tangential port 250 drive flow 251 cylinder (compressor)
252 piston (compressor)
253 Connecting rod (compressor)
254 Crankshaft (compressor)
256 suction valve 257 discharge valve 258 drive gear 259 driven gear 401 driven vehicle 402 driven vehicle 403 transmission medium 405 drive gear 406 driven gear 407 driven gear 407 cam shaft 408 cam 410 combustion chamber 420 spherical cavity 421 valve recess 471 valve cylinder 472 valve piston 473 spring 531 Nozzle 532 Annular Chamber 533 Housing
534 piston 535 spring 536 flange 551 reed valve 552 lead 555 lift check valve 556 spring 557 disc 581 cooler 810 reference profile 811 cam profile 821 hydraulic relay (first hydraulic)
821 Hydraulic relay route (second hydraulic pressure)
821 Hydraulic relay route (third hydraulic pressure)
826 Hydraulic relay 851 Drive 852 Drive 853 Transmission medium
871 oil tank 875 check valve 911 cathode 912 anode 913 DC power supply 914 electrolyzer 915 electrolyte 916 sensor (liquid level, pressure)
917 pump 918 control valve (acid hydrogen)
921 Ultrasonic wave generator 922 High frequency generator

Claims (4)

  1. シリンダヘッドに吸気弁と排気弁を設けた２サイクル内燃機関において、
    排気量より大きい容量の掃気を供給できる掃気供給手段を備え、
    前記掃気供給手段は、
    前記内燃機関にて駆動する圧縮機と、
    吸気通路に掃気増幅手段を設け、
    前記掃気増幅手段は、逆止弁と前記逆止弁の下流に設けた空気流量増幅器から成り、
    前記空気流量増幅器の駆動流通路を前記圧縮機の吐出口に連通し、
    更に、
    燃焼室を略球面状または略円錐状とし、
    前記燃焼室に放射状に吸気弁と排気弁を配置し、
    点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、
    吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポートとし、
    水素、メタンのように空気より密度が小さい燃料、または前記空気より密度が小さい燃料と火花点火式内燃機関またはディーゼル機関の燃料を吸気系統および/または前記燃焼室に供給し、前記内燃機関の運転状況に応じて前記燃料の供給を制御することを特徴とする２サイクル内燃機関。     In a two-stroke internal combustion engine provided with an intake valve and an exhaust valve in a cylinder head,
    It has a scavenging air supply means capable of supplying scavenging air with a capacity larger than the displacement.
    The scavenging air supply means
    A compressor driven by the internal combustion engine;
    Providing scavenging amplification means in the intake passage,
    The scavenging air amplification means comprises a check valve and an air flow amplifier provided downstream of the check valve,
    The drive flow passage of the air flow amplifier is in communication with the discharge port of the compressor,
    Furthermore,
    The combustion chamber is generally spherical or conical,
    Intake valves and exhaust valves are arranged radially in the combustion chamber,
    An ignition plug or an injector is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber,
    The intake port is a tangential port that generates swirl in the cylinder,
    A fuel having a density lower than air, such as hydrogen or methane, or a fuel having a density lower than the air and a spark ignition internal combustion engine or a diesel engine fuel is supplied to an intake system and / or the combustion chamber to operate the internal combustion engine A two-stroke internal combustion engine that controls the fuel supply according to the situation.
  2. 前記燃焼室に放射状に複数の前記吸気弁と排気弁を交互に配置し、
    クランク軸の回転数と同じ回転数で連動する平行な２本のカム軸を設け、
    前記排気弁を前記２本の各カム軸に設けたカムにより開閉し、
    前記吸気弁を逆止弁とする、または、前記吸気弁を前記カムに連動する油圧手段により開閉するまたは前記吸気弁を前記カムとは別のカムに連動する油圧手段により開閉することを特徴とする請求項１に記載の２サイクル内燃機関。     A plurality of the intake and exhaust valves are alternately arranged radially in the combustion chamber,
    Two parallel camshafts interlocking at the same rotational speed as the rotational speed of the crankshaft
    The exhaust valve is opened and closed by a cam provided on each of the two camshafts,
    The intake valve may be a check valve, or the intake valve may be opened and closed by hydraulic means interlocking with the cam, or the intake valve may be opened and closed by hydraulic means interlocking to a cam other than the cam. The two-stroke internal combustion engine according to claim 1.
  3. 前記内燃機関に電気的手段により運転する酸水素発生装置を設け、前記酸水素発生装置で発生する水素または酸水素を、前記空気より密度が小さい燃料として供給することを特徴とする請求項１または請求項２に記載の２サイクル内燃機関。 The internal combustion engine is provided with an oxyhydrogen generator operated by an electrical means, and hydrogen or oxyhydrogen generated by the oxyhydrogen generator is supplied as a fuel having a density smaller than that of the air. The two-stroke internal combustion engine according to claim 2.
  4. ４サイクル内燃機関において、
    燃焼室を略球面状または略円錐状とし、
    前記燃焼室に放射状に複数の吸気弁と排気弁を交互に配置し、
    点火プラグまたはインジェクタを前記燃焼室のシリンダ軸との交点近傍に設け、
    吸気ポートをシリンダ内にスワールを発生させるタンゼンシャルポートとし、
    水素、メタンのように空気より密度が小さい燃料、または前記空気より密度が小さい燃料と火花点火式内燃機関またはディーゼル機関の燃料を吸気系統および/または前記燃焼室に供給し、
    前記内燃機関の運転状況に応じて前記燃料の供給を制御し、
    更に、クランク軸の回転数の１／２の回転数で連動する平行な２本のカム軸を設け、
    前記排気弁を前記２本の各カム軸に設けたカムにより開閉し、
    前記吸気弁を前記カムに連動する油圧手段により開閉するまたは前記吸気弁を前記カムとは別のカムに連動する油圧手段により開閉することを特徴とする４サイクル内燃機関。     In a four-stroke internal combustion engine,
    The combustion chamber is generally spherical or conical,
    A plurality of intake valves and exhaust valves are alternately arranged radially in the combustion chamber,
    An ignition plug or an injector is provided in the vicinity of the intersection with the cylinder axis of the combustion chamber,
    The intake port is a tangential port that generates swirl in the cylinder,
    Supplying a fuel having a density lower than that of air, such as hydrogen or methane, or a fuel having a density lower than that of air and a fuel of a spark ignition type internal combustion engine or a diesel engine to an intake system and / or the combustion chamber;
    Controlling the supply of the fuel according to the operating condition of the internal combustion engine;
    Furthermore, two parallel camshafts interlocking at a rotational speed that is half the rotational speed of the crankshaft are provided,
    The exhaust valve is opened and closed by a cam provided on each of the two camshafts,
    A four-stroke internal combustion engine, wherein the intake valve is opened and closed by hydraulic means interlocking with the cam, or the intake valve is opened and closed by hydraulic means interlocking to a cam different from the cam.
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