WO2019082403A1 - Moteur à combustion interne - Google Patents

Moteur à combustion interne

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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
English (en)
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/fr

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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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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Oil, Petroleum & Natural Gas (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)
  • Valve Device For Special Equipments (AREA)
  • Combustion Methods Of Internal-Combustion Engines (AREA)
  • Combined Controls Of Internal Combustion Engines (AREA)

Abstract

Le problème décrit par la présente invention est qu'un moteur à combustion interne classique a un problème en ce qu'il est impossible de collecter efficacement et de manière fiable un carburant au niveau d'une bougie d'allumage dans une chambre de combustion et de former une couche combustible dans laquelle la propagation de flamme devient uniforme dans la direction circonférentielle, ce qui rend difficile l'amélioration de l'efficacité de combustion. La solution selon l'invention porte sur un moteur à combustion interne à deux temps dans lequel : une chambre de combustion est formée pour avoir une surface sensiblement sphérique ; une soupape d'admission et une soupape d'échappement sont disposées radialement dans la chambre de combustion ; une bougie d'allumage ou un injecteur est disposé à proximité d'un point d'intersection avec un axe de cylindre de la chambre de combustion ; un orifice d'admission est formé d'un orifice tangentiel ; un combustible, tel que l'hydrogène, ayant une densité inférieure à celle de l'air, est alimenté ; et un moyen d'alimentation en air de balayage qui peut fournir de l'air de balayage d'un volume plus grand qu'un volume de déplacement, est en outre prévu, ce qui permet de déplacer le combustible, tel que l'hydrogène, vers l'axe du cylindre par l'action de la centrifugation provoquée par un fort tourbillonnement poursuivi jusqu'à la fin d'une course de compression et d'effectuer une combustion stratifiée. L'invention concerne un moteur à combustion interne à deux temps ou à quatre temps qui est pourvu d'un moyen de combustion dans lequel la soupape d'échappement est ouverte et fermée par des cames disposées sur deux arbres à cames et la soupape d'admission est ouverte et fermée par un mécanisme hydraulique.
PCT/JP2018/000151 2017-10-24 2018-01-09 Moteur à combustion interne WO2019082403A1 (fr)

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Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2610425A (en) * 2021-09-06 2023-03-08 Dolphin N2 Ltd Split cycle internal combustion engine and methods of operating a split cycle internal combustion engine
CN115977789A (zh) * 2023-01-10 2023-04-18 北京理工大学 一种点燃式重油发动机燃烧系统
US11708799B1 (en) * 2022-06-16 2023-07-25 Hong Hue Nguyen System and method for producing hydrogen gas to supply internal combustion engines

Families Citing this family (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2020130771A1 (fr) * 2018-12-18 2020-06-25 Bioactivos Y Nutracéuticos De México S.A. De C.V. Dispositif générateur d'hydrogène pour économie de combustible
IT201900023358A1 (it) * 2019-12-09 2021-06-09 Fpt Ind Spa Motore a combustione interna a ciclo separato
JP7236407B2 (ja) * 2020-05-29 2023-03-09 株式会社三井Ｅ＆Ｓマシナリー 水素燃料を用いた内燃機関
EP4001629B1 (fr) * 2020-11-17 2024-06-26 Volvo Truck Corporation Moteur à combustion interne et véhicule

Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS63306226A (ja) * 1987-05-26 1988-12-14 フオルクスウアーゲン・アクチエンゲゼルシヤフト 混合気圧縮式二サイクル内燃機関
JPH03222817A (ja) * 1989-10-12 1991-10-01 Koichiro Takahashi 吸入掃気分離供給装置付２サイクルエンジン
JP6209802B1 (ja) * 2017-07-04 2017-10-11 正裕 井尻 内燃機関

Family Cites Families (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS6297229A (ja) * 1985-10-21 1987-05-06 Nec Kansai Ltd 螢光膜の形成方法
JPH0713464B2 (ja) * 1987-04-02 1995-02-15 トヨタ自動車株式会社 ２サイクル燃料噴射内燃機関
FR2788307B1 (fr) * 1999-01-07 2001-03-09 Daniel Drecq Moteur compresseur a combustion interne a deux ou a quatre temps
JP2001221070A (ja) * 2000-02-04 2001-08-17 Toyota Motor Corp 電磁駆動弁を有する内燃機関

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS63306226A (ja) * 1987-05-26 1988-12-14 フオルクスウアーゲン・アクチエンゲゼルシヤフト 混合気圧縮式二サイクル内燃機関
JPH03222817A (ja) * 1989-10-12 1991-10-01 Koichiro Takahashi 吸入掃気分離供給装置付２サイクルエンジン
JP6209802B1 (ja) * 2017-07-04 2017-10-11 正裕 井尻 内燃機関

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2610425A (en) * 2021-09-06 2023-03-08 Dolphin N2 Ltd Split cycle internal combustion engine and methods of operating a split cycle internal combustion engine
GB2610425B (en) * 2021-09-06 2023-10-04 Dolphin N2 Ltd Split cycle internal combustion engine and methods of operating a split cycle internal combustion engine
US11708799B1 (en) * 2022-06-16 2023-07-25 Hong Hue Nguyen System and method for producing hydrogen gas to supply internal combustion engines
CN115977789A (zh) * 2023-01-10 2023-04-18 北京理工大学 一种点燃式重油发动机燃烧系统

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