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
  • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
  • F02B33/32—Engines with pumps other than of reciprocating-piston type
  • F02B33/42—Engines with pumps other than of reciprocating-piston type with driven apparatus for immediate conversion of combustion gas pressure into pressure of fresh charge, e.g. with cell-type pressure exchangers
• • 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
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
  • F04B45/00—Pumps or pumping installations having flexible working members and specially adapted for elastic fluids
  • F04B45/04—Pumps or pumping installations having flexible working members and specially adapted for elastic fluids having plate-like flexible members, e.g. diaphragms
  • F04B45/053—Pumps having fluid drive
  • F04B45/0536—Pumps having fluid drive the actuating fluid being controlled by one or more 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
  • F02B75/00—Other engines
  • F02B75/02—Engines characterised by their cycles, e.g. six-stroke
  • F02B2075/022—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle
  • F02B2075/025—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle two

Definitions

• the invention relates to a two-stroke internal combustion engine with a number of cylinders in which pressurized working gas can expand to perform work and thereby drive a crankshaft, wherein the gas outlet window of the or each cylinder is connected to an exhaust system via an exhaust channel, and wherein a switchable multi-way switching valve is arranged in the exhaust channel, via which the gas outlet window is connected to both the exhaust system and a fresh gas pressure line.
• Such a two-stroke combustion engine is made from the EP 4 001 609 A1
• the disclosure is fully incorporated by reference.
• the concept described therein in which, during the final phase of the exhaust stroke of the respective cylinder of the two-stroke engine, the gas-side connection between the exhaust port and the exhaust system is interrupted by means of the multi-way switching valve, and instead the exhaust port is supplied with pressurized fresh gas, preferably pure air, via the exhaust port, enables a particularly efficient, low-emission, and resource-saving operating mode of the two-stroke engine.
• pressurized fresh gas preferably pure air
• a two-stroke engine has inherent advantages over four-stroke engines due to lower frictional power losses, and it is characterized by more favorable NVH behavior ("Noise, Vibration, Harshness”, or "noise, vibration, roughness”, the usual summary of audible and perceptible vibrations on machines, for example in the form of structure-borne sound radiation, torque uniformity and the like).
• the present invention aims to further improve a two-stroke internal combustion engine of the type mentioned above.
• improvements to the known design with regard to scavenging, fuel or lubricating oil consumption, and durability are to be specified.
• particularly preferred design features are to be specified that enable vibration-free or vibration-reduced operation of such engines and thus further improve their NVH (noise, vibration, and harshness) characteristics.
• the multi-way switching valve as a controllable rotary valve for operation at a speed that is an integer multiple of the speed of the crankshaft.
• the invention is based on the premise that a particularly low-loss charge exchange during the purging of fresh air into the respective cylinder via the exhaust port can be facilitated by a particularly rapid opening and closing of the flow cross-sections of the channels in the housing. Since the purging phase must occur within comparatively short valve timings—namely, between the closing of the transfer ports and the closing of the exhaust port in the cylinder—the time spent opening and closing should be kept as short as possible, so that the maximum possible time is available in the fully open state of the valve. According to one aspect of the invention, this is to be achieved by operating the multi-way switching valve, designed as a rotary valve, at a higher speed compared to the respective crankshaft.
• another aspect of the invention provides for the valve to operate at a speed that is an integer multiple of the crankshaft speed. This measure results in at least a doubling of the circumferential speed of the rotary valve in the housing compared to a solution where the valve is operated at a speed identical to the crankshaft speed. This has an exceptionally beneficial effect on the charge exchange of all gas flows controlled by the rotary valve from a fluid mechanics perspective, and Additionally, this design creates the possibility, due to the increased time cross-sections, to reduce the rotary valve diameter again if necessary in order to save installation space.
• the invention proposes an internal combustion engine with at least one working cylinder, one piston per cylinder, and one crankshaft, optionally per cylinder, wherein the working cylinder operates on the two-stroke principle and has at least one exhaust port in the cylinder, wherein a controllable rotary valve is arranged near the exhaust port in a housing surrounding it, wherein the rotary valve can control both the gas flow of the exhaust stream and a gas flow of fresh gas flowing to the cylinder in the exhaust channel, wherein the rotary valve rotates at a crankshaft speed increased by an integer multiple during operation of the internal combustion engine.
• the multi-way switching valve could be controlled and driven by an independent actuator, such as an electric actuator.
• the multi-way switching valve is advantageously driven by the crankshaft.
• the multi-way switching valve can be connected to the crankshaft on the drive side via a transmission with a gear ratio of 1:n, where n is an integer.
• the internal combustion engine is designed as a two-cylinder engine, with each of the two cylinders having its own crankshaft, and the crankshafts rotating in opposite directions during engine operation and coupled to each other via a number of gears.
• the two working cylinders can be arranged coaxially with pistons moving in opposite directions, so that the two working cylinders are combined to form a two-cylinder boxer engine.
• the two working cylinders can be arranged in parallel with pistons moving in the same direction, so that the two working cylinders are combined to form a two-cylinder tandem engine.
• the internal combustion engine can also have only one working cylinder, wherein, according to one aspect of the invention, the free inertial forces of the engine piston and crankshaft assembly are balanced by a second crankshaft with a connecting rod and a pivotable substitute mass attached to it.
• the second crankshaft can be driven by a gear set from the engine crankshaft, in particular rotating in the opposite direction and parallel to the crankshaft and intersecting the cylinder axis.
• the fresh air intended for backflushing and for injection into the exhaust port is, in all variants, most preferably supplied under positive pressure, so that reliable backflushing through the exhaust port into the cylinder is ensured even under the prevailing pressure conditions against the exhaust backpressure.
• the fresh air can be pre-compressed or compressed in a suitably selected compressor unit, for example, an external charging device and/or the engine's own crankcase or piston pump, preferably in a mechanically or electrically driven compressor unit or a mechanically driven diaphragm pump.
• a suitably selected compressor unit for example, an external charging device and/or the engine's own crankcase or piston pump, preferably in a mechanically or electrically driven compressor unit or a mechanically driven diaphragm pump.
• an exhaust-driven charging device is provided for compression, in which, in particular, the enthalpy and/or pressure energy carried in the exhaust stream is used to compress the fresh gas.
• Such an exhaust-driven charging device is designed as a diaphragm pump, which is also controlled via the multi-way switching valve and is connected on the secondary side to the multi-way switching valve via the fresh gas overpressure line and on the primary side to the outlet channel.
• the multi-way switching valve is particularly preferably designed as a three-way valve, via which a gas-side connection between the gas outlet window and the primary side of the compressor unit can be established as required.
• the exhaust gas flowing from the cylinder is fed wholly or partially to the primary side of an exhaust gas charging pump in a first stroke phase of the exhaust stroke and further expanded there, performing work.
• the further expanded exhaust gas in the primary side of the exhaust gas charging pump, together with any exhaust gas still present in the cylinder is routed to the exhaust system.
• the energy of the exhaust gas converted into expansion work on the primary side of the exhaust gas charging pump is thereby converted wholly or partially into compression work on the secondary side of the fresh gas intended for injection into the exhaust port.
• a suitable valve arrangement directs the exhaust gas flow in the exhaust port exclusively to the primary side of the charging device or the exhaust gas charging pump at the beginning of the exhaust opening in the working cylinder.
• both the port connection from the exhaust port to the exhaust system and the port connection from the exhaust gas charging pump to the exhaust system are opened.
• both the exhaust tract to the exhaust system and the exhaust tract to the charging device are closed, and simultaneously the channel for fresh gas injection into the exhaust port is opened.
• the second phase is modified in that the port connection between the exhaust turbine and the The exhaust port or exhaust system remains permanently closed until the first phase of the subsequent cycle begins again.
• the exhaust-side connection of the turbocharger via the valve assembly is only opened to the exhaust port during the phase between "opening the exhaust” and “opening the transfer ports” (the so-called “pre-exhaust”). Any exhaust gas diversion by bypassing the turbocharger would, at most, serve the purpose of limiting boost pressure via a wastegate valve.
• a generator unit alternatively also referred to as a "unit power plant,” comprising a two-stroke combustion engine of the type described above, is considered to have an independent inventive step, as it includes a power generator driven by the two-stroke combustion engine.
• the ensemble-like combination of a two-stroke combustion engine of the type described above with an associated power generator driven by it is considered to have an independent inventive step.
• both the two-stroke combustion engine and the power generator can be installed underground, which is considered to have an independent inventive step.
• Such a concept of housing energy generation or supply components underground is considered to have an independent inventive step in view of the advantages it offers. with regard to the avoidance of noise emissions, improved sound insulation and the like, are considered to be fundamentally independent and inventive.
• One aspect of the invention involves enclosing the entire motor and/or generator unit with a multi-part sound-absorbing material, advantageously constructed from multiple layers of materials with different sound absorption coefficients.
• a horizontal division of the multi-part sound-absorbing housing with a removable top cover for easy maintenance of the unit proves advantageous.
• the housing preferably has openings for the intake and exhaust of fresh air to the motor, as well as cooling air inlets and outlets for the generator and motor. Furthermore, cooling water lines for heat dissipation or heat recovery and electrical wiring can be routed through the housing.
• media and/or electrical lines can be provided and routed through the housing, to which a heat exchanger can be suitably connected.
• This heat exchanger allows the exhaust heat to be extracted from the engine or exhaust system and transferred to a cooling medium.
• the lines through the housing are preferably located in one of the housing's partition planes.
• a further embodiment provides for the complete unit to be arranged underground in a watertight enclosure, the enclosure advantageously sealed at the top with a walkable cover.
• the enclosure can also be lined with sound-absorbing material.
• the engine's intake and exhaust air can be routed to the surface via ducts and can extend as pipes or pipe bundles above the roofs of nearby buildings to prevent disturbances from exhaust air and noise emissions in residential areas.
• the engine-generator unit can be fixed in the enclosure by vibration-damping bearings, such as rubber dampers or hydraulic mounts, or it can be suspended freely on springs, with the springs attached to the enclosure or a subframe.
• the underground housing of the engine-generator unit described here which is considered to be an independent inventive, not only offers the advantage of soundproofing but also reduces the space required by above-ground installations. Furthermore, this protects Housing the unit protects it from direct environmental influences and damage, helps keep the engine at operating temperature more easily during the cold season, and certainly has aesthetic advantages due to its invisibility.
• a vehicle in particular a two-wheeler, a passenger vehicle, or a commercial vehicle, with a two-stroke combustion engine of the type described above is also considered to be independently inventive.
• this vehicle can further comprise a power generator driven by the two-stroke combustion engine, which in turn is connected to a storage battery at its output.
• a sodium-ion battery is particularly preferred and considered an independent inventive embodiment for the energy storage device.
• the motor for the application of the motor as a range extender/hybrid with a driven generator, it is thus possible, according to an independent inventive aspect, to combine the motor-generator unit with a sodium-ion battery instead of the commonly used lithium-ion battery.
• the advantages of the sodium-ion battery lie in its approximately 40% lower manufacturing costs compared to the lithium-ion battery in terms of storage capacity, its non-flammability, and the readily available raw material worldwide.
• the disadvantage of the sodium-ion battery achieving only about 40% capacity per unit volume, is not significant, as the battery capacity can be considerably reduced in combination with the onboard generator.
• a battery capacity of 5-10 kWh is quite sufficient, which corresponds to a reduction in battery capacity of approximately 5-30% compared to purely battery-electric vehicles. This ratio can, of course, also be applied to larger vehicle classes.
• the control strategy of this hybrid vehicle powered by a range extender is not to drive as long as possible on battery power alone, but to activate the range extender regularly at intervals of 10-30 minutes depending on the driving profile.
• the range extender At full load, the range extender would operate continuously. This control strategy allows engine waste heat to be used almost continuously for heating purposes, especially if an additional, highly insulated coolant heat storage tank is provided as a thermal battery.
• the range extender operates advantageously at a fuel-consumption-optimized full-load point or within a narrow full-load speed range for power generation.
• the engine's design as a "true" boxer engine (cylinder axes are arranged coaxially) is advantageous for vehicle applications, as it is completely free of inertial forces and moments and can be built with a very flat profile.
• the significantly smaller battery allows the space freed up in the floor panel to accommodate the boxer engine, whose flat design is thus very well suited to the space constraints in the vehicle floor of battery-electric vehicles.
• the engine and its tank system can be designed for different fuels (natural gas, LPG, ethanol, e-fuels, gasoline, etc.) for which a sufficiently comprehensive network of filling stations exists, and in addition, multifuel designs with several separate tank and fuel systems for the engine can be provided to be more flexible with regard to different fuels.
• fuels natural gas, LPG, ethanol, e-fuels, gasoline, etc.
• This two-stroke engine with its scavenging process and fresh gas charging can, in principle, be combined with all variants of an electric drive (wheel hub motors, central motor), especially in the form of a generator drive, such as Plug-in Hybrid Vehicle (PHEV), Hybrid Vehicle (HEV), battery electric vehicle with range extender (BEV-Rex).
• PHEV Plug-in Hybrid Vehicle
• HEV Hybrid Vehicle
• BEV-Rex battery electric vehicle with range extender
• a heat pump with a compressor that can be driven by a two-stroke combustion engine of the type described above is also considered to be independently inventive.
• aspects of the invention thus relate in particular to an internal combustion engine, which is preferably intended for use as a range extender in vehicles, as a drive for compressor heat pumps and power generators, preferably in close proximity to buildings.
• the application area close to buildings places particularly high demands on the engine's NVH (Noise, Vibration, Harshness) behavior.
• NVH Noise, Vibration, Harshness
• compatibility with different synthetic and mineral-based fuels is required, as Alcohols, natural gas, liquefied petroleum gas (LPG), hydrogen, ammonia, gasoline, diesel fuel, and future synthetic fuels (e-fuels) are all potential fuels.
• LPG liquefied petroleum gas
• e-fuels future synthetic fuels
• Power generators located near buildings could become increasingly important in the future, as they can be used as charging stations for battery-powered electric vehicles (EVs), preferably when powered by synthetic fuels.
• EVs battery-powered electric vehicles
• a suitable power generator for this purpose must have excellent noise and vibration (NVH) characteristics to avoid disturbing residents' sleep and an electrical output of approximately 5-25 kW, depending on the size and number of vehicles to be charged.
• NSH noise and vibration
• the waste heat from the combustion engine should ideally be able to be used for heating and hot water production in the surrounding buildings (combined heat and power).
• the generator can also serve as a grid-supplementary or grid-independent power source for driving an electric motor-driven heat pump.
• the variable-speed generator can also regulate the speed of the heat pump drive via the adjustable AC or three-phase frequency, thus controlling the desired heat pump output.
• an exhaust gas recirculation pump is mounted on the engine for compressing fresh gas, its exhaust line (pulse line) can advantageously be connected directly to the rotary valve.
• the pulse line or the exhaust cover of the exhaust gas recirculation pump can be designed as a heat exchanger to dissipate thermal energy from the system.
• a temperature gradient occurs in the exhaust gas recirculation pump due to expansion in the exhaust chamber, so that the exhaust gas mass flow mixed in the rotary valve from the engine cylinder and the exhaust gas recirculation pump still has a temperature level of approximately 250–450°C and is preferably routed directly after the rotary valve and before the exhaust silencer through a catalyst for exhaust gas treatment.
• this heat exchanger is preferably positioned in the exhaust pipe after the catalytic converter and before the muffler.
• the heat energy extracted from the exhaust gas via the heat exchanger can either be used as thermal energy for heating purposes or to power a steam process via an expansion engine (turbine, piston engine, etc.), which performs technical work and is, for example, mechanically coupled to the engine's crankshaft.
• liquid water can be added to the gas mixture in the working cylinder separately or directly to the fuel.
• Such water injection particularly into the cylinder of a turbocharged two-stroke engine, can be advantageously used to reduce NOx formation by lowering the temperature during combustion, and/or to shift the knock limit to higher compression ratios and increase the expansion work by evaporating water to increase cylinder pressure after top dead center.
• turbocharged engines that drive the compressor using exhaust gas energy (exhaust gas pump, turbocharger)
• the additional water vapor content in the exhaust gas mass flow provides the turbocharger with a higher exhaust gas enthalpy, which, in an independently inventive embodiment, can be used to increase boost pressure.
• a water injection valve is positioned close to the cylinder in one of the transfer ports, preferably with its injection direction towards the cylinder to achieve minimal wall wetting in the transfer port, the water is introduced into the cylinder during the scavenging phase. In this position, the injection valve is permanently exposed to low pressures and temperatures, which, according to one aspect of the invention, results in a simple and cost-effective design.
• the water injection during the scavenging phase with the onset of evaporation, advantageously cools the piston crown and reduces the temperature of the gas mixture at the beginning of the compression stroke via the enthalpy of vaporization of the injected water.
• the water should be introduced near top dead center (TDC), in accordance with an inventive aspect of the invention.
• TDC near top dead center
• the evaporation of the water, and thus its expansion, does not occur significantly during the compression phase, so that the piston does not have to perform any or only minimal compression work.
• the pressure increase associated with evaporation directly benefits the indicated piston work.
• the introduction of water near TDC can be achieved in the two-stroke engine described here by means of high-pressure injection with an injector positioned in the cylinder head that injects directly into the combustion chamber.
• This design allows for free selection of the injection timing, with the option of advancing the injection timing even before the ignition timing to achieve a compromise between high expansion work and the reduction of nitrogen oxide formation.
• installing a water injection nozzle in the cylinder head presents no problems; however, this design is the most expensive because the nozzle must operate under high temperatures and water pressures.
• a compromise between the additional financial outlay for a Water injection and the resulting increase in expansion work via the engine piston result from the independently inventive possibility of placing a water injection nozzle in the cylinder wall, near the cylinder bore, with the spray direction into the cylinder, wherein the cylinder is connected to the nozzle of the injection valve via a small bore or a shot channel.
• the position of this shot channel can advantageously be located between the "exhaust port closes" position and top dead center.
• the advantage of this arrangement lies in the shielding of the injection valve by the piston when the highest combustion temperatures and pressures prevail at top dead center.
• the variation of the injection timing is limited, since injection cannot occur later than when the piston overflows the shot channel.
• the particularly preferred position of the shot channel is thus chosen as a compromise between the desired effects of nitrogen oxide reduction or knock limit shift and increased expansion work.
• the amount of water supplied to the cylinder for each power stroke can be used, in an aspect also considered independently inventive, to precisely control the temperature rise of the working gas during the compression phase for each individual power stroke.
• This and advantageously in conjunction with a variable compression ratio of the engine, allows the necessary ignition conditions for a compression ignition process based on homogeneous charge compression ignition (HCCI) to be achieved. This not only increases the thermodynamic efficiency of the engine but also reduces emissions such as CO, HC, and NOx.
• HCCI homogeneous charge compression ignition
• One control strategy for HCCI combustion is the combination of a load-dependent change in the compression ratio, whereby high compression ratios are set at low loads and lower values are set as the load increases. Since the variation in the compression ratio, i.e., the change in combustion chamber volume, is mechanically limited and usually occurs “slowly" in relation to changes in engine load and speed, the compression ratio can be chosen so high that, depending on fuel type, engine temperature, air-fuel ratio, etc., the required final compression temperature for auto-ignition of the combustion gas is reliably reached before the thermodynamically ideal ignition point. The resulting premature shift in the ignition point, caused by the excessively high final compression temperature, is then corrected by the cycle-specific addition (injection) of water, with its temperature-reducing evaporative cooling, to the desired temperature. Ignition timing is controlled. Further parameters for controlling HCCI combustion can include combustion chamber temperature and crankshaft unevenness, the measured values of which can be detected by sensors.
• the HCCI combustion process can be combined with the spark ignition process using existing spark plugs or pre-chamber spark plugs in the engine's combustion chamber.
• Compression ignition is primarily used in the low-load range, while spark ignition is used in the high-load range when precise control of the ignition timing is no longer possible.
• compression-ignition four-stroke engines generally exhibited poor control over the desired ignition timing at high loads.
• the control variable was often the amount of uncooled exhaust gas added to regulate the temperature of the working gas. Even small fluctuations in the exhaust gas volume resulted in significant deviations in the temperature rise during the compression phase, leading to either misfires or premature ignition with undesirable pressure increases.
• This ignition sensitivity of the four-stroke compression engine is due, among other things, to the high volumetric efficiency under full load and the excellent ignition properties of a stoichiometric fuel-air mixture, resulting from the catalytic three-way exhaust aftertreatment system.
• the two-stroke engine has lower delivery rates per working cycle and higher exhaust gas content, which, in combination with a lean fuel-air mixture, leads to slower ignition behavior, thus making it possible, according to one aspect of the invention, to extend compression ignition to higher loads.
• gaseous hydrogen (H2) or oxyhydrogen (HHO) can be provided for the aforementioned internal combustion engine, but also, in principle, for other internal combustion engines.
• gaseous hydrogen or oxyhydrogen can be added to the gas mixture in the working cylinder separately before its ignition or directly to the fuel. Both improve the ignitability of the fuel-air mixture.
• This can be advantageously used for both spark-ignition combustion and the described compression ignition, in particular by adding H2 or HHO to fuel-air mixtures under partial load with a high residual gas content, or by improving the ignition of very lean fuel-air mixtures under high load.
• the hydrogen is advantageously added in gaseous form close to the cylinder during the purging phase in order to obtain a clear quantity measurement for each individual work cycle.
• the hydrogen can be injected directly into the cylinder via a shot channel.
• the production and supply of H2 or HHO can be carried out directly at the engine during operation by means of an electrically driven electrolysis system that draws electrical energy from the vehicle's electrical system.
• the quantities of hydrogen to be stored are very small, as they are produced in real time according to the engine's operating point.
• the amount of hydrogen added to the fuel for ignition improvement is less than 10% by mass.
• the water required for the splitting process is either provided in a separate refillable tank or can also be obtained on-board by condensing the engine exhaust gas.
• the in Fig. 1 The schematically depicted internal combustion engine 1 is of a design according to the EP 4 001 609 A1 It is designed as a two-stroke engine. It comprises two cylinders, of which in Fig. 1 Only one is shown, and each contains a working piston 4.
• the working piston 4 acts on a crankshaft 8 via a connecting rod 6.
• the working piston(s) 4 of several or all cylinders 2 can also act on a common crankshaft 8.
• the working chamber 10 is located within cylinder 2, where a compressed fuel-air mixture is combusted during the power stroke of cylinder 2.
• the working piston 4 which is slidably arranged within cylinder 2, executes a power stroke, thereby driving the crankshaft 8.
• the combustion gas is fed as exhaust gas to an exhaust system 12 connected to the exhaust side of cylinder 2 during an exhaust stroke.
• the working chamber 10 is connected on the gas inlet side via a number of gas inlet windows 14 opening into the working chamber 10 to the inlet transfer ports, and on the exhaust side via an exhaust port 18 or an exhaust port system connected to the exhaust system 12 via an exhaust port 18 or an exhaust port system connected to the gas outlet window(s) 16 opening into the working chamber 10.
• the control of the gas exchanges in the working chamber 10 is effected in the manner of a conventional two-stroke engine by the up and down movement of the The working piston 4 in the cylinder opens or closes the gas inlet port 14 and the gas outlet port 16 as appropriate.
• the gas outlet port 16 opens first during the expansion stroke, thus opening the exhaust port to reduce the cylinder pressure by releasing exhaust gas (necessary pre-exhaust). Subsequently, the gas inlet port 14 opens, and with it the transfer ports, which introduce fresh gas from the crankcase (or turbocharger) into cylinder 2.
• the type of cylinder scavenging (reverse flow, parallel flow, etc.) is irrelevant, as all methods aim to expel the exhaust gas in cylinder 2 with the incoming fresh gas from cylinder 2 with minimal mixing of the two gas phases.
• the exhaust port typically remains open longer than the intake transfer ports. During this phase towards the end of the charge cycle, the undesirably high proportions of fresh gas scavenging losses occur, subsequently entering the exhaust port 18 and from there the exhaust system 12.
• the two-stroke combustion engine 1 has a switchable multi-way switching valve 30 located in the exhaust port 18, through which the gas outlet port 16 is connected to both the exhaust system 12 and a fresh gas pressure line 32.
• the multi-way switching valve 30 allows for an alternative gas-side connection between the gas outlet port 16 and the exhaust system 12 or between the gas outlet port 16 and the fresh gas pressure line 32.
• the multi-way switching valve 30 shuts off the gas flow from cylinder 2 towards the exhaust system 12. Simultaneously, by closing the exhaust port 18 to the exhaust system 12, the multi-way switching valve 30 opens the fresh gas pressure line 32 to the exhaust port 18 at cylinder 2.
• This valve position is in Fig. 1 As shown. In this valve position, pre-compressed fresh gas, preferably pure pre-compressed air, is forced into the exhaust port 18. This fresh gas pushes the gas mass present in the exhaust port 18, i.e., the exhaust gas with the contained fresh gas scavenging losses including fuel, back into cylinder 2 and simultaneously ensures a boost or charging in cylinder 2 at the end of the charge cycle.
• the timing design of the port windows in the cylinder and the overall design of the two-stroke combustion engine are preferred. 1. Smaller channel volumes support high specific engine outputs, while larger volumes reduce scavenging losses.
• the volume of the section of the exhaust channel 18 located between the gas outlet window 16 of cylinder 2 and the multi-way switching valve 30 should preferably be at least 10% and/or at most 40% of the geometric displacement volume of cylinder 2.
• the fresh gas pressure line 32 is, in a configuration considered to be independently inventive, connected at its inlet side to the secondary side of an exhaust gas-driven charging device 40, which thus forms a compressor unit for the fresh gas pressure line 32.
• the exhaust gas-driven charging device 40 is also configured, in a configuration considered to be independently inventive, as an exhaust gas charging pump, such as those found, for example, in the EP 2 846 019 A1 , from the EP 2 846 020 A1 , from the EP 3 061 970 A1 or from the EP 3 282 109 A1 is known. It includes a diaphragm pump 42, which is connected on the secondary side via the fresh gas overpressure line 32 to the multi-way switching valve 30 and on the primary side to the outlet channel 18.
• the multi-way switching valve 30 is designed as a three-way valve, via which a gas-side connection between the gas outlet window 16 and the primary side of the diaphragm pump 42 can be established if required.
• FIG. 2 Figure 1 shows the two-stroke internal combustion engine 1 during the exhaust stroke at the end of the first phase, which serves for exhaust gas charging, shortly before the gas intake ports 14 open.
• first phase of exhaust gas charging when the gas exhaust port 16 opens, the connection to the exhaust system 12 is just barely completely closed by means of the valve position of the multi-way switching valve 30, and the exhaust gas is fed exclusively to the primary side of the diaphragm pump 42 of the exhaust gas charging pump via the valve assembly, in order to expand there by converting the kinetic energy of the exhaust gas.
• This first phase of the charge exchange corresponds in function and timing approximately to the "pre-exhaust" of a conventional port-controlled two-stroke engine.
• the timing at which the exhaust port 18 opens via the The piston edge is exclusively connected to the exhaust gas charging pump (exhaust gas charging) and is preferably between approximately 10-50°KW, whereas the connection from the exhaust port 18 to the exhaust system 12 via the multi-way switching valve 30 is released approximately 0-25°KW before the gas inlet windows 14 open.
• the multi-way switching valve 30 has already opened the flow path from the outlet channel 18 to the exhaust system 12 and opened the flow path of the relaxed, reverse exhaust mass flow from the diaphragm pump 42 of the exhaust charging pump to the exhaust system 12.
• the primary side of the exhaust charging pump is preferably permanently connected to the exhaust system 12 to allow for the longest possible period of exhaust gas relaxation in the exhaust charging pump 42. This corresponds to the operating mode generally intended for the exhaust charging pump 42.
• the two-stroke combustion engine 1 is in a third or final phase of the charge exchange, also corresponding to the situation in Fig. 1
• this final phase when the gas outlet window 16 in cylinder 2 begins to close due to the upward-moving working piston 4, the multi-way switching valve 30 closes both the flow paths between the exhaust port 18 and the exhaust system 12, and the one between the exhaust port 18 and the diaphragm pump 42 of the exhaust gas charging pump.
• the multi-way switching valve 30 opens a flow path that now connects the exhaust port 18 to the fresh gas pressure line 32 and thus to the secondary or fresh air side of the exhaust gas charging pump.
• the gas outlet window 16 is closed by the upward-moving working piston 4, thus completing the charge exchange.
• several cylinders 2 can be supplied by an exhaust gas charging pump 42 via one or more such valve devices 30.
• the multi-way switching valve 30 can generally be designed as a device with oscillating and/or rotating valves for controlling the gas exchange in the outlet channel 18.
• it is designed as a roller valve, in particular as a cylindrical, axially rotatable rotary valve 44 in a stationary control housing 46.
• the control housing 46 can be formed in one piece with the associated cylinder 2, which is easily achievable using casting techniques.
• the present embodiment is designed for a further improvement in efficiency and a further reduction in scavenging losses.
• the rotary valve 44 of the two-stroke internal combustion engine 1 is designed for operation at a speed that is an integer multiple of the speed of the crankshaft 8, namely, in a particularly preferred embodiment, at twice the speed of the crankshaft 8.
• the control housing 46 is enlarged in perspective view in Fig. 5 shown.
• the Figures 6 to 8 the associated rotary valve 44, designed as a roller valve, in perspective views ( Figs. 6 and 7 ) and in cross-section ( Fig. 8 ).
• the control housing 46 advantageously and according to one aspect of the invention, being constructed in one piece as a monolithic block or in the manner of an integrated design as part of the cylinder 2, has a central receiving channel 48 with a cylindrical inner cross-section into which the roller- or cylinder-shaped rotary valve 44 can be inserted. Together, the rotary valve 44 and the control housing 46 thus combined form the multi-way switching valve 30.
• the control housing 46 has, as shown in Fig. 5 It can be seen that there is an outlet opening 50 that can be connected to the silencer of the exhaust system 12, a channel 52 that can be connected to the exhaust turbocharger and a channel 57 that is connected to the outlet channel 18.
• the channel elements 54 opening onto the inner wall of the receiving channel 48 are connected to the fresh gas overpressure line 32.
• the channel segments 56 arranged together with the outer surface of the rotary valve body 44, connect the switchable gas channels 54 to a channel 57, depending on the rotational position of the rotary valve 44 relative to the control housing 46, in order to push fresh gas into the outlet channel 18.
• the channel segments 56 in the rotary valve 44 are responsible for the fresh gas flow, with a further channel segment 58 provided for the exhaust gas flow.
• the base body of the rotary valve 44 forms the control edges 60.
• the rotary valve 44 traverses the corresponding channels 50, 52, 54, 57 in the control housing 46 to carry out the gas exchange in accordance with the concept of the invention.
• the rotary valve 44 is provided with control cutouts that correspond to channel windows in the control housing 46.
• the aforementioned doubling of the rotational speed of the rotary valve 44 relative to the crankshaft speed is made possible in the exemplary embodiment by a suitable selection and design of the pairings of the channel elements 50, 52, 54, 57 in the control housing 46 and the channel segments 56, 58 on the roller body of the rotary valve 44.
• the channel elements 54 are positioned in the control housing 46 in a plurality of longitudinally offset tracks 61a,b, wherein, in the exemplary embodiment, each track 61a,b performs the same specific switching function for fresh gas control depending on the rotational position of the rotary valve 44.
• a desired switching function which also enables the synchronous control of the gas flows even at as explained above, the increased rotational speed of the rotary valve 44 is achieved by the suitable arrangement of the functionally different channel routing of several tracks 61a,b.
• the tracks 61a,b in the control housing 46 refer to the following in the illustration. Fig. 5
• the opening cross-sections for the compressed fresh gas for backflushing via the outlet channel 18 into the cylinder 2 correspond at the rotary valve 44 with the channel segments (56a and 56b) designed as channel pockets, which are joined in the area 56c to push compressed fresh gas via the channel 62 in the control housing 46 into the outlet channel 18.
• This concept of the correspondingly designed tracks 61 is considered to be independent and autonomously inventive.
• FIG. 9 The assembled multi-way diverter valve 30, i.e., with the rotary valve 44 inserted into the receiving channel 48, is shown in longitudinal section.
• the channel segments 56, 58 of the rotary valve 44 preferably seal against the channel elements 50, 52, 54, 57 in the control housing 46 without contact, and according to an independent inventive aspect, via a gap 64 which is in the range of 3 to 25 hundredths of a millimeter.
• the rotary valve 44 and/or control housing 46 can be provided with coatings 65 that keep the sealing gap height small and are break-in capable.
• sliding varnishes or plastics can be provided as soft coatings that are partially worn away during the break-in phase without causing consequential damage, thus creating a minimal sealing gap between the rotary valve 44 and the control housing 46.
• the control housing 46 is equipped on its cylindrical inner surface facing the rotary valve 44, forming the receiving channel 48, with a soft coating 65a of the type mentioned, whereas the rotary valve 44 may preferably have a resistant hard coating 65b, which may, for example, be a nickel or chrome coating.
• the effective gap height between rotary valve 44 and control housing 46 depends, among other things, on the thermal expansion of the two components, which are preferably designed to be liquid-cooled.
• the cooling of the rotating rotary valve 44 as shown in the illustration in Fig. 9 is also clearly discernible, with a The coaxial inlet and outlet (66, 68) of coolant are arranged counter-currently concentric to the axis of rotation 70. This arrangement allows the coolant flow in the rotary valve 44 to be realized in a space-saving manner with only one seal 72 at one shaft end 74.
• the rotary valve 44 is advantageously designed, and according to one aspect of the invention, as already mentioned above, as a roller valve, with channel segments 56, 58 incorporated on its outer circumference for exhaust gas and fresh gas flow.
• the rotary valve is preferably mounted coaxially on both sides of the gas exchange control area in rolling bearings 78, which preferably seal off the respective shaft bearing 76 on both sides via radial shaft seals, each bearing 78 having an inlet and outlet for the lubricant from the engine's oil sump.
• a simplified, also preferred and independently inventive embodiment of the rotary valve mounting alternatively provides closed rolling bearings 78 with sealing discs and grease filling, as is the case in Fig. 9 As shown.
• sliding sealing rings 80 preferably gap-sealing and non-contacting, can be arranged between the pressurized gas-carrying control area of the rotary valve 44 and the bearing points/shaft seals to create a pressure drop in front of the sealing or bearing points.
• the leakage gas between the sliding sealing ring 80 and the rolling bearing 78 is advantageously fed to the engine's intake air via bypass channels.
• the intended use of the engine concept described herein as a drive source for power generators, heat pumps, or vehicles places not only low NVH (noise, vibration, and harshness) but also particularly high ecological and economic demands on specific fuel consumption, exhaust emissions, low maintenance, and durability.
• NVH noise, vibration, and harshness
• the invention provides a design solution that reduces the two-stroke engine's inherent oil loss in lubrication to the necessary minimum and also enables virtually maintenance-free operation and high durability for all tribologically stressed engine components.
• the loss oil lubrication according to one aspect of the invention is applied in particular only to the lubrication of the connecting rod bearings 90, 98 and the cylinder bore.
• FIG. 10a Figure 1b schematically shows the two-stroke internal combustion engine 1 in high section with separate oil bath 82 for the crankshaft main bearings 81 and the drive gears 84, with separate supply 86 of loss oil through one of the crankshaft journals to the crankshaft crankpin 88 of the connecting rod bearing 90, with separate loss oil supply 94 directly to the cylinder wall for lubricating the working piston 4, and further with bores 96 for separate loss oil supply for the piston pin 98 or for the upper connecting rod bearing 100.
• Fig. 10a Figure 1b schematically shows the two-stroke internal combustion engine 1 in high section with separate oil bath 82 for the crankshaft main bearings 81 and the drive gears 84, with separate supply 86 of loss oil through one of the crankshaft journals to the crankshaft crankpin 88 of the connecting rod bearing 90, with separate loss oil supply 94 directly to the cylinder wall for lubricating the working piston 4, and further with bores 96 for separate loss oil supply for the piston pin 98 or for the upper connecting rod bearing 100
• FIG. 10b shows a vertical section through the crankshaft axis with a one-piece assembly of cylinder 2 and rotary valve housing 46, the oil supply 94 for piston lubrication and the oil-air blowing 112-120 for piston pin lubrication.
• Fig. 10c In contrast, the figure shows the working piston 4 of the two-stroke internal combustion engine 1 in a horizontal section through the piston pin axis with the bores 96 for supplying air-oil mixture to the piston pin connecting rod bearing 98.
• FIG. 10a shows, in particular, a cross-section through the cylinder axis, that the two crankshaft main bearings 81 are designed as rolling bearings located in the oil bath 82.
• the oil bath 82 is sealed off from the interior of the crankcase by shaft seals 102 between the crankshaft main bearings 81 and the crank discs.
• the crankshaft main bearings 81 no longer require lubrication via a loss-and-loss oil system.
• the connecting rod bearing 90 of the crankshaft crankpin 88 is supplied directly with metered loss-and-loss oil via channels machined into the crankshaft 8, preferably as shown in Figure 1.
• the oil exiting the crankshaft journal 88 is flung into the crankcase during operation, where it mixes with the intake fresh air.
• the connecting rod bearing 100 on the piston pin 98 is lubricated by the oil-air mixture in the crankcase when this mixture is directed past the piston interior and into the engine cylinder's scavenging ports during the purge phase.
• the lubricating oil supply to the piston pin bearing 88 can be achieved through at least one window 108 in the piston skirt approximately at the level of the piston pin 98. Pre-compressed fresh gas from the crankcase is directed through this window into the transfer port area when the port is open, as is the case in... Fig. 10b as is evident.
• FIG 10a , b Figure 1 shows a pressure line 112 branching off from the interior of the crankcase 110, which is preferably arranged at the lowest gravimetric point (depending on the installation position of the engine 1) of the interior of the crankcase 110 in order to carry away any residual oil from the crankcase 110 in the gas phase.
• the pressure line 112 is connected via a check valve 114 to a pressure accumulator 116, from which a pressure line 118 branches off, opening into the cylinder interior (also referred to above as the working chamber 10) at an inlet 120 at the lower end of the cylinder bore.
• the inlet 120 is preferably located approximately at the level of the piston pin 88 when the piston 4 reaches its bottom dead center.
• metered oil can be added to the pressure line 118 near the cylinder wall via a supply line 122.
• crankcase 110 When the engine 1 is started, pressure fluctuations occur inside the crankcase 110 due to the volume change of the oscillating piston 4. In each cycle, fresh gas is compressed in the crankcase 110 as the piston 4 moves towards bottom dead center. Shortly before the piston 4 opens the scavenging ports 14 to the working chamber 10 ( Fig. 10b , position 124), the highest compression pressure in the crankcase 110 is reached. Up to this point, fresh gas from the crankcase 110, possibly containing oil, is pumped via the check valve 114 to the pressure accumulator 116, where it remains, since the piston skirt closes the opening of the pressure line 118 into the working chamber 10 at the feed point 120, thus preventing a short-circuit flow inside the crankcase 110.
• fresh gas-oil blowing can be applied to the bearing of the piston pin 98 from both sides, as is done in Fig. 10a based on the two feed-in points 120 and shown there Fig. 10c as illustrated by the two radial bores 96 shown there.
• the pressure accumulator 116 including its lines, preferably has a volume of 2-10% of the stroke volume of the working cylinder 2.
• FIG. 10b Figure 1 shows the supply of oil to the cylinder wall through corresponding channels and lubrication pockets in the cylinder bore below and opposite the exhaust port 18, with each channel near the cylinder bore being equipped with a check valve 126 to prevent backflow of oil due to pressure fluctuations in the crankcase.
• the oil supply is located below the exhaust ports or on the opposite side of the cylinder bore, since the piston skirt bears against these areas due to the pivoting movement of the connecting rod under load.
• one or more cylinders 2 of the internal combustion engine 1 can, for example, be designed for the addition of water to reduce NOx emissions (achievable, for example, by lowering the temperature during combustion), to shift the knock limit to higher compression ratios, and/or to increase the expansion work of the working piston 4.
• this can be achieved by adding liquid water to the gas mixture in the working cylinder or directly to the fuel.
• the additional water vapor content in the exhaust gas mass flow can provide the turbocharger with a higher exhaust gas enthalpy, which, in an independently inventive embodiment, can be used to increase the boost pressure.
• such a continuous supply of water to the motor 1, which is considered advantageous in stationary applications, such as Charging stations for electric vehicles, power generators, or heat pump drives can be supplied via existing water pipe networks near buildings, and also via the condensation of water vapor from engine exhaust gases.
• another aspect of the invention offers a very simple solution: adding water to the fuel in a desired ratio. This can be done with water-miscible fuels such as methanol, ethanol, and ammonia.
• a water injection valve can be arranged in the housing of the cylinder 2 in positions 184a, b, c close to the cylinder 2 in or at one of the transfer ports (position 184c), its injection direction, defined by its longitudinal axis, preferably being oriented towards the cylinder 2.
• This along with suitable control, allows the wall wetting in the transfer port to be kept particularly low by introducing water into the cylinder 2 during the scavenging phase.
• the water injection valve is permanently exposed only to comparatively low pressures and temperatures, which, according to one aspect of the invention, enables a simple and cost-effective design.
• the indicated mean effective pressure can be increased particularly significantly via the piston movement by the amount of water introduced into cylinder 2.
• the amount of water can be introduced into cylinder 2 near top dead center. The evaporation of the water, and thus its expansion, does not occur during the compression phase, so that no increased compression work is required from the working piston 4. The The pressure increase associated with evaporation thus directly benefits the indicated piston work.
• water injection near top dead center can be achieved in the two-stroke internal combustion engine 1 described herein by means of high-pressure injection with a water injection valve 185, which is positioned in position 184a in the cylinder head 186 and injects directly into the combustion or working chamber 10.
• a water injection valve 185 which is positioned in position 184a in the cylinder head 186 and injects directly into the combustion or working chamber 10.
• a compromise between the additional financial costs of water injection and the gain in expansion work via the working piston 4 results from the independently inventive possibility of placing a water injection nozzle 185 in position 184b in the cylinder wall, near the cylinder bore, with the injection direction into the cylinder 2, wherein the cylinder 2 is connected to the nozzle of the injection valve 185 via a small bore or a shot channel 188.
• the position of this shot channel 188 can advantageously be located between the "exhaust port closes" position and top dead center.
• the advantage of this arrangement lies in the shielding of the injection valve 185 by the working piston 4 when the highest combustion temperatures and pressures prevail at top dead center.
• the variation of the injection timing is limited, since injection cannot occur later than when the working piston 4 passes over the shot channel 188.
• the particularly preferred position 184b of the firing channel 188 thus results as a compromise between the desired effects of nitrogen oxide reduction or knock limit shift and increased expansion work.
• the amount of water supplied to cylinder 2 for each working stroke can be used, in an aspect also considered independently inventive, to precisely control the temperature rise of the working gas during the compression phase by means of a temperature reduction via enthalpy of vaporization for each individual working stroke.
• This and advantageously in conjunction with a variable compression ratio of the engine, allows the necessary ignition conditions for a compression ignition process based on homogeneous charge compression ignition (HCCI) to be achieved, which not only improves the thermodynamic efficiency of the engine.
• HCCI homogeneous charge compression ignition
• the compression ignition not only increases the efficiency of the engine 1, but also reduces emissions such as CO, HC and NOx.
• FIG. 12 Figure 1 shows an example of a two-stroke cylinder head 190 in which a variable compression ratio is enabled by an axially displaceable combustion chamber insert 192.
• This insert seals the combustion chamber 195 against a fixed squish insert 196 with a squish gap 198 via sealing rings 194.
• the turbulence-promoting squish gap 198 which is formed by the piston in a position near top dead center and the squish insert 196, remains unchanged.
• a preferred control strategy for HCCI combustion is the combination of a load-dependent change in the compression ratio, whereby high compression ratios are set at low load and lower values are set with increasing load. Since the variation of the compression ratio, i.e., the change in the combustion chamber volume, is usually mechanically limited and occurs relatively slowly in relation to changes in engine load and speed, the compression ratio can be selected so high that, depending on the fuel type, engine temperature, air-fuel ratio, etc., the required final compression temperature for auto-ignition of the combustion gas is reliably reached before the thermodynamically ideal ignition point.
• the resulting premature ignition point, caused by the excessively high final compression temperature, is then controlled by the cycle-specific addition (injection) of water with its temperature-reducing evaporative cooling to the desired ignition point.
• Other parameters for controlling HCCI combustion can include combustion chamber temperature and crankshaft irregularity, the measured values of which can be detected by sensors.
• aspects of the invention further include a new structural and tribological design of the crankshaft bearings, the coupling of crankshafts 8 to one another, the drive of the rotary valves 44 via the crankshaft(s) 8, and the drive of mass balancing gears for vibration damping.
• the crankshaft main bearings 81 are tribologically separated from the interior of the crankcase 110 (crankcase pump) by placing the necessary shaft seal 102, which seals the interior of the crankcase 110 against the atmosphere, between the crankshaft flange and the bearing. 81st place (see also Fig.
• crankshaft main bearings 81 are separated from the respective interiors of the corresponding crankcases 110.
• the crankshaft main bearings 81 are designed as rolling bearings, which in turn are advantageously supplied with lubricant via an oil bath 82 or spray lubrication.
• a plain bearing or an oil circulation lubrication system, or a combination of these designs can also be provided.
• the rotary valves are also mounted with rolling bearings, which are advantageously supplied with lubricant from the same oil bath 82 as the crankshaft bearings, as shown, for example, in the illustrations in Fig. 13 , 14 removable.
• These show the respective gear cascade, which distributes the oil from the oil bath 82 through the meshing of some gears 216 in the oil bath 82 in the enclosed space of the gear drive.
• the rotary valve 44 can be driven electrically, i.e., by means of an associated electric motor, and thus be mechanically separate from the crankshaft drive.
• phase adjustment between the rotary valve 44 and the crankshaft position by means of suitable control of such an electric drive can also be provided in order to adapt the charge exchange, which can be influenced by the valve device via variable phase, in particular with regard to the timing of the fresh air scavenging via the exhaust port 18, to the load and speed of the engine 1.
• Individual control cutouts in the control housing 46 can also be individually variably designed with respect to flow cross-section and phase.
• the multi-way switching valve is driven directly or indirectly by the crankshaft 8.
• the multi-way switching valve 30 designed as a controllable rotary valve 44, at a speed that is an integer multiple of, preferably twice, the speed of the crankshaft 8, as provided for in one aspect of the invention, the multi-way switching valve 30 can be connected to the crankshaft 8 on the drive side via a transmission 210, according to aspects of the invention.
• the transmission ratio of the transmission 210 is 1:n, where n is an integer natural number.
• crankshafts 8 (one for each cylinder 2) are provided, arranged parallel to each other, their axes of rotation intersecting the common cylinder axis 214, and advantageously coupled to each other at the same rotational speed in opposite directions via a gear pair 216.
• the gears are spur gears 218, each of which is assigned to one of the rotary valves 44, and each of which is coupled to the gear pair 216 via an intermediate gear 220 to form the transmission gear 210.
• the transmission ratio of the transmission gear 210 is set to 1:2, according to the design in the exemplary embodiment.
• An alternative design for a two-cylinder engine is the arrangement of the two cylinder units (cylinders with crankshaft drive) as a so-called tandem engine 230, as described in Fig. 14 This is shown. It has two axially parallel, counter-rotating crankshafts 8 of the same speed, which are advantageously – analogous to the above-described design of the boxer engine 212 – as in Fig. 14a
• the cylinders are shown to be coupled via gears 232 of a gear pair 216.
• the cylinder axes are parallel to each other and point in the same direction, with both cylinders 2 lying in a plane perpendicular to the axes of the two crankshafts 8.
• the pistons move synchronously "up and down” in their respective cylinders 2.
• the free second-order inertial forces can be controlled, according to an aspect considered to be independently inventive, by the The two opposing rotary valves 44 and their double crankshaft speed are balanced if they have a corresponding imbalance and act as a Lancaster compensating gear.
• Deviations from the ideal cylinder arrangements of the "real" boxer or tandem engine 212, 230 described here could cause vibrations due to free forces or moments, which, up to a certain level of preferably about 15%, may be considered acceptable in practice.
• both cylinders 2 operate in phase, i.e., they also fire together.
• the two crankcases are hermetically connected and equipped with a common throttle valve for fresh gas supply/regulation. This achieves a uniform delivery rate in the common crankcase pump and avoids synchronization problems with multiple throttle valves. It is also possible to provide a common muffler, catalytic converter, and exhaust gas charging pump for both cylinders 2 to reduce installation space and costs.
• the described two-cylinder arrangements can also be multiplied and combined, for example, to form four-, six-, or eight-cylinder engines.
• the cylinder base housing 234 is preferably formed from a cylinder base housing 234 usable for both cylinders 2, which accommodates a cylinder block 238 on parallel opposing parting planes 236 that intersect the crankshaft main bearings 81 centrally.
• the cylinder block 238 also includes the receptacle for the rotary valve 44.
• the crankshaft assembly including the engine pistons, can be pre-assembled and inserted into the cylinder 2 from below and then closed with the cylinder base housing 234.
• the two cylinders 2 can be bolted together with a screw thread passing through the cylinder base 234 or each can be bolted separately into the cylinder base 234.
• the throttle valve on the cylinder base housing 234 supplies the fuel.
• both crankcases are also connected inside the crankcase 110 and are supplied with fresh air via a common throttle valve.
• a virtually vibration-free design can be achieved, according to an aspect considered independently inventive, by installing a second (balance) crankshaft parallel to the engine crankshaft in the engine cylinder axis, as in the two-cylinder boxer engine 212.
• This second crankshaft has a corresponding mass weight at the end of the small connecting rod eye, also guided in the cylinder axis via the connecting rod.
• the mass weight either performs an oscillating motion or, alternatively, is mounted laterally and performs a pivoting motion, with its center of gravity located in or near the cylinder axis.
• the stroke of this (balance) crankshaft can be smaller than that of the engine crankshaft, in conjunction with an increase in mass at the end of the small connecting rod eye.
• the 1:2 speed ratio of the respective crankshaft 8 to the rotary valve 44 can be adjusted, according to an independently inventive aspect, by directly driving the respective rotary valve 44 via the respective crankshaft 8.
• the drive considered particularly advantageous, can be implemented using only a single gear pair 216 or alternatively via one or more intermediate gear sets 220.
• the twice-as-rotary speed of the rotary valve 44 relative to the crankshaft 8 can be used to compensate for second-order free forces from the crank mechanism by providing the rotary valve 44 with a suitable imbalance. This can be implemented particularly advantageously in the arrangement as a tandem two-cylinder engine 230, in which, as in Fig.
• the illustration shows two counter-rotating, opposing rotary valves 44, one for each cylinder 2.
• This type of mass balancing also called Lancaster mass balancing, can also be applied to other shafts that may be present, provided they have twice the crankshaft speed and are arranged symmetrically to a plane in which the center of gravity of both pistons moves.
• the spur gears 218 are advantageously lubricated, as shown, from the oil bath 82 shared with the crankshaft main bearings 81, preferably by immersion lubrication of at least one gear. If two or more separately parallel crankshafts 8 are provided, as in the "true" boxer engine 212 described above or the tandem engine 230, the two crankshafts 8 are preferably coupled via the same gears 218 that are also responsible for driving the respective rotary valve 44.
• the torque output of the entire engine is advantageously provided on the crankshaft side of the gear drive, preferably at one of the crankshaft ends.
• the spur gears 218 can also be helical.
• a mass balancer If, as is often desirable in single-cylinder engines, a mass balancer is provided, it can also be driven by the spur gear 218 of the respective crankshaft 8 and supplied with lubricant by the common oil bath 82.
• the complete gear drive including the crankshaft main bearings 81 and rotary valve bearings, runs in a volume separate from the interior of the crankcase 110 and is preferably also sealed oil-tight to the outside from the atmosphere.
• the oil sump can be designed to be maintenance-free according to one aspect of the invention, by using a so-called lifetime oil fill.
• Immersion lubrication preferably achieved by meshing or immersing the tooth flanks in the oil bath 82 and the subsequent distribution of lubricating oil by centrifugal forces on rotating components, can be advantageously used to supply lubricant to bearing points in the form of oil collection pockets at bearing points inside the housing, which allow the oil to flow to the desired lubrication points via channels and gravity.
• each of the two cylinders 2 is thus assigned a crankshaft 8, wherein the crankshafts 8 rotate in opposite directions to each other during operation of the engine 1 and are coupled to each other via a number of gears 216.
• a generator block 240 also referred to as a "block power plant"
• a generator block 240 is considered to be independently inventive and includes such a two-stroke internal combustion engine 1, which is connected via a drive shaft 242 to a power generator 244 and drives the latter during operation.
• Fig. 15 An embodiment considered to be independently inventive is shown in which the power generator 244 is electrically connected on its output side to a wall-box 246 for charging an electric vehicle 248 near a building 250.
• the generator 244 of the combined heat and power plant 240 is electrically connected on its output side to a heat pump 252 installed in building 250, which is intended for heating purposes.
• a heat pump 252 installed in building 250
• these two aspects can also be combined; that is, the generator 244 can be connected on its output side to both the wallbox 246 and the heat pump 252, whereby the output line can be appropriately controlled and distributed to these components as needed.
• the generator 244 is also designed for particularly favorable NVH (noise, vibration, and harshness) characteristics in order not to disturb the residents' sleep.
• NVH noise, vibration, and harshness
• the waste heat from the combustion engine 1 can advantageously be supplied to the surrounding buildings 250 for heating purposes or for hot water preparation (combined heat and power).
• the generator in addition to using the generator as a charging station for an electric vehicle 248, it can also serve as a grid-supplementary or grid-independent power source for driving the heat pump 252.
• the speed-controlled generator 244 can also regulate the speed of the electric motor 272 of the heat pump drive, particularly via the adjustable AC or three-phase frequency, and thus regulate the desired heat pump output in a speed-controlled manner.
• the waste heat from the combustion engine 1 is preferably supplied to the refrigeration circuit of the heat pump 252 between the evaporator 255 and the compressor via a heat exchanger 253.
• the thermal energy of the engine exhaust gases is extracted from the exhaust gas via a heat exchanger 330 between the rotary valve 44 or catalyst 332 and the silencer 12.
• this can be achieved down to a temperature level below the condensation temperature of the water content.
• the power plant 240 and in particular its components two-stroke combustion engine 1 and power generator 244, is installed underground and preferably also encapsulated, in a further development considered to be independently inventive.
• Such a concept of housing energy generation or supply components in underground construction is considered to be fundamentally independent inventive in view of the advantages achievable with regard to the avoidance of noise emissions, improved sound insulation, and the like.
• the invention provides, in one aspect, for enclosing the entire motor and/or generator unit with a multi-part soundproof shell 254 made of sound-absorbing material.
• This shell can advantageously be made of multiple layers of materials with different sound absorption coefficients.
• a horizontal division 256 of the soundproof shell 254, forming a removable upper cover 258, is advantageously provided, allowing for particularly easy maintenance of the internally arranged power unit 240 and its components.
• the soundproof housing 254 is preferably equipped with openings for the fresh air supply and exhaust gas discharge of the combustion engine 1, or with cooling air inlets and outlets for the generator 244 and combustion engine 1. Furthermore, cooling water lines for heat dissipation or heat recovery and electrical wiring can be routed through the housing 254. These routings through the housing 254 are preferably located in one of the housing partition levels 256.
• the complete power plant 240 can be arranged underground in a watertight basin 260, the basin 260 advantageously being closed at the top with a walkable cover.
• the basin 260 can additionally be lined with sound-absorbing material.
• the intake and exhaust air of the combustion engine 1 can be routed to the above-ground atmosphere via ducts 262. These can be designed in a chimney-like manner as pipes 264 or pipe bundles extending above the roof 266 of nearby buildings 250 in order to prevent disturbances from exhaust air and noise emissions in residential areas.
• the motor-generator unit 240 can be fixed in the housing 260 by vibration-damping bearings, such as rubber dampers or hydraulic mounts, or it can be suspended freely on springs, which may be attached to the housing 260 or an auxiliary frame.
• the underground housing of the motor-generator unit 240 described here which is considered to be an independent inventive, not only offers the advantage of soundproofing but also reduces the space required by above-ground installations. Furthermore, this housing protects the unit from direct environmental influences and damage, helps maintain the combustion engine 1 at operating temperature more easily during the colder months, and its concealment certainly offers aesthetic advantages.
• Fig. 17 The concept of a heat pump 252, whose compressor 270 can be driven by a two-stroke combustion engine 1 of the type described, was also considered.
• this drive is implemented as an indirect drive, namely via the current generator 244.
• a drive can also be provided directly, as shown in the illustration.
• Fig. 17 can be removed.
• the power generator 244 can, in a particularly simple design, simply be replaced by the compressor 270.
• the underground arrangement considered particularly advantageous and independently inventive, can be retained.
• the compressor 270 of the heat pump 252 and the power generator 244 can both be provided together in the underground arrangement, preferably, for example, by mechanical switching, they could be operated alternately.
• the motor 1 can be operated during the day as a drive for the heat pump 252 and at night as a drive for the power generator 244, for example for charging the electric vehicle.
• a vehicle 280 In the manner of a possible and preferred mobile use of the two-stroke internal combustion engine 1, a vehicle 280, considered to be independently inventive, is equipped with such a two-stroke internal combustion engine 1.
• the vehicle 280 is in Fig. 18a shown as an example in a perspective top view of the chassis 282 of a four-wheeled passenger vehicle; of course, the construction concept described below is also applicable to two-wheelers or commercial vehicles.
• Fig. 18b shows the chassis 282 of vehicle 280 in a side view, Fig. 18c front view and Fig. 18d Top view.
• the two-stroke internal combustion engine 1 designed in this case as a boxer engine 212 of the type described above, is mounted on the chassis 282 near the rear axle 284 of the vehicle 280.
• a current generator 244 which can be driven by the two-stroke internal combustion engine 1 according to a further development considered to be independently inventive, is arranged directly on the boxer engine 212.
• This generator is electrically connected on its output side to a battery pack 286 designed as an energy storage device.
• the battery pack 286 comprises four storage batteries mounted directly on the chassis 282, each designed as a sodium-ion battery 288 in this exemplary embodiment.
• an electric drive motor 290 is provided for the actual propulsion of the vehicle 280.
• This motor is electrically connected to the battery pack 286 on the input side, thus drawing its energy from the battery pack, and its output side transmits power to the front axle 294 via a differential gear 292.
• a fuel tank 296 for the fuel for the two-stroke combustion engine 1 a water tank 298 for storing injection water for the water injection system of the type described above, and a water tank 300 for storing hot water heated by the waste heat of the two-stroke combustion engine 1 are mounted on the chassis 282.
• the combination, as envisaged according to this aspect of the invention, of the comparatively inexpensive two-stroke engine 1 for on-board power generation, in conjunction with a relatively small storage battery in the form of the battery pack 286, and the vehicle drive via at least one electric motor as the "actual" drive motor 290, enables a very cost-effective and energy-efficient drive concept that combines all the advantages of electric drive with the energy density of liquid or gaseous fuels and, in the future, enables CO2-free emissions via e-fuels.
• the drive axles can, of course, be freely selected in this concept; that is, either the front axle 294 or the rear axle 284, or both axles 284, 294 (all-wheel drive), can be equipped with electric drive motors.
• the placement of the two-stroke combustion engine 1 near the rear axle 284 is not mandatory but can, in principle, be freely chosen within the vehicle 280.
• engine designs based on the tandem cylinder arrangement are also feasible in the vehicle, as are inline engines and V-engines.
• the production and supply of H2 or HHO can be carried out directly at the two-stroke combustion engine 1 during operation by means of an associated, electrically driven electrolysis system 302, which draws electrical energy from the vehicle's electrical system 314.
• the quantities of hydrogen to be stored are very small, as they are produced in real time according to the engine's operating point.
• the quantities of hydrogen to be supplied to the fuel for ignition improvement are less than 10% by mass.
• the water required for the splitting process is either supplied in the separate refillable water tank 298 or can also be obtained on-board by condensing the engine exhaust gas.
• a correspondingly designed device for hydrogen/oxyhydrogen gas supply is shown schematically in Fig. 19 shown, with two possibilities for introducing hydrogen into the working space 10 of the cylinder 2 being illustrated.
• the two-stroke internal combustion engine 1 is shown in the scope of its partially depicted cylinder 2, the working piston 4 running therein, the connecting rod 6, and the crankshaft 8.
• a metering valve 312 is located upstream of the working chamber 10 of the cylinder 2 on the gas side and is connected on the gas-side inlet side to the electrolysis system 302.
• This system is electrically connected on the inlet side to the symbolically depicted vehicle electrical system 314 of the vehicle 280 and, on the media-side, is connected via lines (not shown) to the water tank 298 or a [unclear text] for the supply of the water to be split.
• the system is connected to an alternative water supply.
• the metering valve 312 allows for the sequential or continuous supply of H2 or HHO into the overflow channels 14 for mixing with fresh gas from the crankcase 110.
• the low-pressure alternative for adding hydrogen to the fresh gas refers to the injection position 318 near the throttle valve in the intake airflow. Position 318 can be located either directly after or directly before the throttle valve. Hydrogen is preferably supplied continuously at this point. Control can be achieved via a metering valve 316 or directly via the power supply to the electrolysis unit 302.
• FIG. 20 The schematic shows the embodiment of the two-stroke combustion engine 1, which is considered to be independently inventive, with exhaust-side catalyst 320 and heat exchanger 322, for the removal of thermal exhaust energy, as well as with end silencer 12.

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• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Chemical & Material Sciences (AREA)
• Combustion & Propulsion (AREA)
• Output Control And Ontrol Of Special Type Engine (AREA)

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Citations (7)

• 2024
  • 2024-07-22 EP EP24190145.3A patent/EP4685343A1/fr active Pending
• 2025
  • 2025-07-16 WO PCT/EP2025/070410 patent/WO2026021979A1/fr active Pending

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