Two-stroke internal combustion engine and use of the two-stroke internal combustion engine

Abstract

A two-stroke internal combustion engine (1) having a number of cylinders (2) in which pressurized working gas can work-performingly expand, as a result of which a crankshaft (8) is driven, wherein the gas outlet window (16) in the or each cylinder (2) is connected to an exhaust system (12) via an outlet duct (18), and wherein a switchable multi-way switching valve (30), via which the gas outlet window (16) is connected both to the exhaust system (12) and to a fresh-gas overpressure line (32), is located in the outlet duct (18), is intended to be still further improved in particular in terms of scavenging, fuel or lubricant consumption and durability. For this purpose, according to the invention, the multi-way switching valve (30) is designed as a controllable rotary valve (44) for operation at a rotational speed of an integer multiple of the rotational speed of the crankshaft (8).

Description

[0001] Description

[0002] Two-stroke internal combustion engine and use of the two-stroke internal combustion engine

[0003] 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.

[0004] Such a two-stroke combustion engine is known from EP 4 001 609 A1, the disclosure of which is incorporated in its entirety ("incorporation 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. This is based on the understanding that a major reason for the typically adverse exhaust gas or emission values ​​of conventional two-stroke engines is the scavenging losses during gas exchange and the associated comparatively low capture efficiency.In general, 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).

[0005] Specifically, EP 4 001 609 A1 discloses a slot-controlled two-stroke piston engine which allows fresh gas recirculation via a switchable exhaust port into the cylinder and also utilizes this concept for fresh gas charging. The charge exchange for this purpose occurs particularly in conjunction with an exhaust gas charging pump acting as a compressor. This concept reduces the disadvantages of a two-stroke engine described above, resulting from scavenging losses, and thus allows the fundamental advantages of the two-stroke engine to be utilized.

[0006] The present invention aims to further improve a two-stroke internal combustion engine of the type mentioned above. In particular, improvements to the known design with regard to scavenging, fuel or lubricating oil consumption, and durability are to be specified. Furthermore, 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.

[0007] With regard to the two-stroke internal combustion engine of the type mentioned above, this problem is solved according to the invention by designing 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.

[0008] Advantageous embodiments of the invention are the subject of the dependent claims.

[0009] 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 control times—namely, between the closing of the transfer ports and the closing of the exhaust port in the cylinder—the time spent on each opening and closing should be kept as short as possible, so that the maximum possible time can be spent 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 rotational speed compared to the respective crankshaft.To ensure the timing accuracy of the aforementioned purging process with respect to the engine cycle, one 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 within the housing compared to a solution where the valve operates at a speed identical to the crankshaft speed. This has an exceptionally advantageous effect on the fluid dynamics of the charge exchange of all gas flows controlled by the rotary valve. Furthermore, this design allows for a reduction in the rotary valve diameter, if necessary, due to the increased cross-sectional area, thus saving installation space.

[0010] In other words, according to this aspect, 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.

[0011] The multi-way switching valve could be controlled and driven by an independent actuator, such as an electric actuator. However, for a particularly compact and cost-effective design, the multi-way switching valve is advantageously driven by the crankshaft. In a particularly advantageous embodiment, 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. Most preferably, the multi-way switching valve is designed for operation at twice the speed of the crankshaft; this is achievable with a gear ratio of 1:2, i.e., n = 2.

[0012] In a particularly preferred embodiment, the transmission is designed as a gear drive, the gears and, if applicable, the shaft bearings for lubrication are connected to a common oil circuit. On the one hand, the individual gears can be connected to a common oil circuit for lubrication. On the other hand, and for the sake of a particularly simple design and efficient lubrication, the gears can advantageously also be connected to a common oil circuit together with the crankshaft main bearing(s). In a preferred embodiment, the oil circuit can include an oil bath. Oil circulation can be achieved, in a preferred embodiment, by meshing at least one gear through the oil bath and / or by immersing at least one gear in the oil bath.

[0013] In a particularly advantageous embodiment, the internal combustion engine is designed as a two-cylinder engine, with each of the two cylinders having its own crankshaft. During engine operation, the crankshafts rotate in opposite directions and are coupled to each other via a number of gears. In one embodiment, the two working cylinders can be arranged coaxially with pistons moving in opposite directions, thus forming a two-cylinder boxer engine. Alternatively, in another embodiment, the two working cylinders can be arranged in parallel with pistons moving in the same direction, thus forming a two-cylinder tandem engine. However, all configurations of a two-cylinder engine as a V-engine are also possible, with cylinder bank angles between 0° and 180°.

[0014] In an alternative embodiment, considered to be independently inventive, 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.

[0015] 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. To enable this, 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.In a particularly preferred embodiment, which is considered to be independently inventive, an exhaust gas-operated charging device is provided for compression, in which, in particular, the enthalpy and / or pressure energy carried in the exhaust gas stream is used to compress the fresh gas.

[0016] Particularly preferred, and in an embodiment considered to be independently inventive, a two-stroke engine designed in this way is combined with an exhaust gas charging pump for the purpose of compressing the fresh gas, such as those known, for example, from EP 2 846 019 A1, EP 2 846 020 A1, EP 3 061 970 A1, or EP 3282 109 A1. The disclosures contained in these documents are expressly and fully incorporated by reference. In a further advantageous embodiment, and for the sake of a particularly efficient design, such an exhaust gas-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 exhaust port.The multi-way switching valve is particularly preferred as a three-way valve, which allows a gas-side connection between the gas outlet window and the primary side of the compressor unit to be established as required.

[0017] In this particularly preferred embodiment, which is considered to be independently inventive, during the exhaust stroke, 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. In a second stroke phase of the exhaust stroke, 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. Advantageously, 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.

[0018] Advantageously, in a first phase of the exhaust stroke, 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. In a second phase, 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. In a third phase, during the closing of the exhaust port in the working cylinder, 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. In an alternative advantageous embodiment, particularly when using a turbocharger as the charging device, the second phase is modified in that the port connection between the exhaust turbine and the exhaust port is opened.The exhaust system remains permanently closed until the first phase of the subsequent cycle begins again. In other words, 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.

[0019] In each of the embodiments considered to be independently inventive, the internal combustion engine is used according to aspects of the invention: for driving an electric generator, preferably in a stationary arrangement in an underground position, in a mobile arrangement in a vehicle, preferably for driving a generator carried mobile in the vehicle, wherein in a mobile vehicle application the generator is preferably coupled with a storage battery, preferably designed as a sodium-ion battery, as an energy storage device; for directly driving the compressor of a heat pump; for directly driving an electric heat pump independently of the grid, in that the internal combustion engine drives a generator, the generated electrical energy of which is supplied to the electric drive motor of the heat pump, wherein a change in the speed of the internal combustion engine or the generator is advantageously achieved.the generator with accompanying frequency change can be used to change the speed of the electric drive of the heat pump compressor; for heating purposes, by using the waste heat of the motor wholly or partially for heating purposes, in particular by feeding the motor waste heat into the refrigeration circuit after an evaporator and before a compressor.

[0020] Accordingly, a generator unit, alternatively also referred to as a "block 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. In other words, 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. In a design considered to have an independent inventive step, both the two-stroke combustion engine and the power generator can be installed underground.Such a concept of housing components for energy generation or supply in underground construction is considered to be fundamentally inventive in terms of the advantages achievable with regard to the avoidance of noise emissions, improved sound insulation and the like.

[0021] In such a design, sound insulation is of particular importance, especially the soundproofing of the engine in conjunction with an electric generator, which is intended to run overnight near buildings. The exhaust gas charging pump(s), as one of the primary sources of structure-borne noise, is advantageously integrated completely into the fresh air filter housing(s) or intake air silencer housing(s) of the engine to reduce its noise emissions. This allows the noise emissions from the exhaust gas charging pump housing covers to be dampened together with the incoming intake air. For a two-cylinder engine, either two exhaust gas charging pumps are used, or, in the case of two cylinders operating in phase, a single exhaust gas charging pump with a correspondingly larger airflow would also be applicable.One aspect of the invention involves enclosing the entire engine and / or generator unit with a multi-part sound-absorbing material, which is advantageously made of multiple layers of materials with different sound absorption coefficients. A horizontal division plane in 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 fresh air supply and exhaust gas discharge of the engine, as well as cooling air inlets and outlets for the generator and engine. Furthermore, cooling water lines for heat dissipation or heat recovery and electrical wiring can be routed through the housing. Alternatively or additionally, media and / or electrical lines can be provided and routed through the housing, connecting to a heat exchanger through which the exhaust gas heat from the engine or generator is transferred.The exhaust gas stream can be extracted and fed into a cooling medium, provided it is suitably connected. For the sake of a particularly simple design, the pipe routing through the housing is preferably located in one of the housing's partition planes.

[0022] 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. According to aspects of the invention, 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 motor-generator unit described here, which is considered an inventive design in its own right, 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 engine at operating temperature more easily during the colder months, and its concealment certainly offers aesthetic advantages.

[0023] Similarly, 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. According to a further development 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. The combination of an inexpensive two-stroke engine for on-board power generation, in conjunction with a small storage battery and the vehicle drive via at least one electric motor, as provided for in this aspect of the invention, enables a very cost-effective 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.

[0024] Particularly preferred, and in a configuration considered to be independently inventive, is the use of a sodium-ion battery as the energy storage device. Specifically, for the application of the motor as a range extender / hybrid with a driven generator, it is thus provided, 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 materials worldwide. The disadvantage of the sodium-ion battery, namely achieving only about 40% capacity per unit volume, is not of significant consequence, since the battery capacity can be considerably reduced in combination with the onboard generator.For small cars, a battery capacity of 5-10 kWh is perfectly adequate, which corresponds to a reduction in battery capacity of approximately 5-30% compared to purely battery-electric vehicles. This ratio can, of course, be applied to larger vehicle classes as well. The control strategy of this hybrid vehicle, equipped with a range extender, is not to drive as long as possible on battery power alone, but rather to activate the range extender at regular intervals of 10-30 minutes, depending on the driving profile.

[0025] 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 mass forces and torques and can be built with a very low profile.Provided that the existing production platforms of battery-electric vehicles can be used cost-effectively for the range-extender drive, the space freed up in the floor plate for the boxer engine can be used due to the significantly smaller battery volume, and its flat design is therefore very well suited to the installation space requirements in the vehicle floor of battery electric vehicles.

[0026] 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.

[0027] This two-stroke engine, with its scavenging process and fresh gas charging, can in principle be combined, particularly in the form of a generator drive, with all variants of an electric drive (wheel hub motors, central motor), such as plug-in hybrid vehicles (PHEVs), hybrid vehicles (HEVs), and battery electric vehicles with range extenders (BEV-Rex). Similarly, 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.

[0028] 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. Furthermore, compatibility with various synthetic and mineral-based fuels is required, such as alcohols, natural gas, liquefied petroleum gas (LPG), hydrogen, ammonia, gasoline, diesel fuel, and future synthetic fuels (e-fuels). A low specific fuel consumption is also a key requirement, which should enable overall efficiencies of over 40%. This presents a particular challenge, as the aforementioned application areas already encompass drive power outputs of 3–30 kW, which partially corresponds to the small engine segment.As is generally known, the percentage of mechanical and thermal losses in an engine increases with decreasing displacement, which is why small engines (q=20-25%) achieve significantly lower overall efficiencies than large engines (q=45-50%). Additionally, the engine waste heat from cylinder and cylinder head cooling and the exhaust gases should be available for further use. This would particularly benefit heating systems in buildings or industrial settings, especially in conjunction with heat pumps, to reduce the temperature difference in such systems by feeding the engine waste heat into the refrigerant after the heat source (evaporator) and before the compressor, thus significantly increasing the heat pump's coefficient of performance (COP).

[0029] 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. A lack of charging options for EVs, such as the scarcity of public charging stations, their unpredictable availability, and the unavailability of solar charging stations during the day at the vehicle's parking location, often makes charging the vehicle battery overnight at home unavoidable. A suitable power generator for this purpose must have excellent noise, vibration, and harshness (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. The waste heat from the combustion engine should ideally be used for heating purposes.The generated energy can be supplied to surrounding buildings for hot water preparation (combined heat and power). In addition to using the generator as a charging station for an electric vehicle, it can also serve as a grid-supplementary or grid-independent power source to drive an electric motor-driven heat pump. When operating an electric motor-driven heat pump independently of the grid, 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.

[0030] If 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.

[0031] If the exhaust gas flow is to be routed through a heat exchanger to recover thermal energy, 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.

[0032] According to an aspect considered to be independently inventive, 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 pollutant 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. In turbocharged engines that drive the compressor using exhaust gas energy (exhaust gas charging 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. All four of the above can be applied to the engine concept described by aspects of the invention.Effects can be achieved through constructive measures. In stationary applications such as electric vehicle charging stations, power generators, or heat pump drives, the continuous supply of water to the engine can be ensured either via existing water supply networks near the building or by condensing water vapor from the engine exhaust gases. In addition to self-contained water injection independent of the fuel used, 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.

[0033] If 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.Both the lower gas temperature of the fuel-air mixture at the start of combustion and the water vapor content, with its considerable heat capacity, significantly reduce the peak temperature during combustion, thus contributing to a substantial reduction in nitrogen oxide formation. The temperature drop during the compression phase generally shifts the knock limit to higher compression ratios, which can be used to increase engine efficiency.

[0034] If the intention is to maximize the indicated mean effective pressure (IEC) via piston movement by means of the amount of water introduced into the cylinder, the water should be introduced near top dead center (TDC), in accordance with an inventive aspect. 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 increased compression work, or only to a very limited extent. The pressure increase associated with evaporation directly benefits the indicated piston work. According to aspects of the invention, 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 reduced nitrogen oxide formation. While accommodating a water injection nozzle in the cylinder head presents no problems in terms of installation space, this design is the most expensive because the injection nozzle must operate under high temperatures and water pressures. A compromise between the additional cost of water injection and the gain in expansion work via the engine piston is achieved by placing a water injection nozzle in the cylinder wall, near the cylinder bore, with the spray direction into the cylinder, and connecting the cylinder to the nozzle of the injection valve via a small bore or a spray channel.The position of this injection channel can advantageously lie 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. However, the variation of the injection timing is limited, since injection cannot occur later than the point at which the piston overflows the injection channel. According to one aspect of the invention, the particularly preferred position of the injection channel is thus chosen as a compromise between the desired effects of nitrogen oxide reduction or knock limit shift and increased expansion work.

[0035] Furthermore, 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. Particularly in two-stroke engines under partial load, with their high residual exhaust gas content in the combustion chamber, compression ignition enables reliable ignition at low loads and with lean fuel-air mixtures.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 load and lower values ​​as the load increases. Since the variation of the compression ratio, i.e., the change in 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 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 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.

[0036] Furthermore, 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.

[0037] In the past, compression-ignition four-stroke engines generally exhibited poor controllability at high loads with regard to the desired ignition timing. 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. The reason for this ignition sensitivity of the four-stroke compression engine lies, among other things, in 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.

[0038] However, the two-stroke engine exhibits lower delivery rates per combustion cycle and higher exhaust gas content, which, in combination with a lean fuel-air mixture, leads to a more sluggish ignition response. According to one aspect of the invention, this makes it possible to extend compression ignition to higher loads. In another aspect, also considered independently inventive, the addition of gaseous hydrogen (H2) or oxyhydrogen (HHO) to the working gas before ignition can be provided for the aforementioned internal combustion engine, but also, in principle, for other internal combustion engines. In this embodiment of the invention, gaseous hydrogen or oxyhydrogen can be added to the gas mixture in the working cylinder separately before ignition or directly to the fuel. Both improve the ignitability of the fuel-air mixture.This fact 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 ignition in very lean fuel-air mixtures at high load. The hydrogen is advantageously added in gaseous form close to the cylinder during the scavenging phase to ensure precise metering for each individual combustion cycle.

[0039] Similar to the water injection described above, the hydrogen can be injected directly into the cylinder via a shot channel.

[0040] According to aspects of the invention and in a concept considered inventive independently of the foregoing, 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 electrolysis process is either provided in a separate, refillable tank or can also be obtained on-board by condensing the engine exhaust gas.

[0041] An embodiment of the invention is explained in more detail with reference to a drawing. The drawing shows:

[0042] FIG. 1 schematically shows a two-stroke internal combustion engine, FIG. 2 schematically shows the two-stroke internal combustion engine according to FIG. 1 during the exhaust stroke in an early, first phase,

[0043] FIG. 3 schematically shows the two-stroke internal combustion engine according to FIG. 1 during the exhaust stroke in a second phase,

[0044] FIG. 4 schematically shows the two-stroke internal combustion engine according to FIG. 1 during the exhaust stroke in a third or final phase,

[0045] FIG. 5 shows the control housing of a multi-way switching valve of the two-stroke internal combustion engine according to FIG. 1 in perspective view.

[0046] FIGS. 6 to 8 show the rotary valve of the multi-way switching valve of the two-stroke internal combustion engine according to FIG. 1 in perspective view (FIGs. 6, 7) and in cross-section (FIG. 8),

[0047] FIG. 9 shows the assembled multi-way switching valve of the two-stroke internal combustion engine according to FIG. 1 in longitudinal section.

[0048] FIG. 10a, 10b schematically show a two-stroke internal combustion engine according to FIG. 1 with a supply of loss oil for the lubrication of the piston or piston pin bearing in a high-section view.

[0049] FIG. 10c shows a piston of the two-stroke internal combustion engine according to FIG. 1 in perspective view,

[0050] FIG. 11 shows a cylinder of the two-stroke internal combustion engine according to FIG. 1 with additional water injection in longitudinal section,

[0051] FIG. 12 shows a two-stroke cylinder head with a variably adjustable compression ratio,

[0052] FIG. 13 shows a boxer engine with (FIG. 13a) and without (FIG. 13b) a reduction gearbox, each in frontal view; FIG. 14 shows a tandem engine with (FIG. 14a) and without (FIG. 14b) a reduction gearbox, each in frontal view.

[0053] FIG. 15 shows a generator block with a two-stroke internal combustion engine according to FIG. 1 in use as a charging station for a battery-electric vehicle,

[0054] FIG. 16 shows a generator block with a two-stroke internal combustion engine according to FIG. 1 in use for driving an electric heat pump with heat input of the engine waste heat into the refrigeration circuit,

[0055] FIG. 17 shows a generator block with a two-stroke internal combustion engine according to

[0056] FIG. 1 in a use for driving an electric generator and a heat pump compressor with heat input of the engine waste heat into the refrigeration circuit of a heat pump,

[0057] FIG. 18 schematically shows a vehicle designed as a hybrid vehicle or with a range extender and a two-stroke internal combustion engine according to FIG. 1 ,

[0058] FIG. 19 schematically shows a two-stroke internal combustion engine according to FIG. 1 in high section with hydrogen or oxyhydrogen gas supply, and

[0059] FIG. 20 schematically shows a two-stroke internal combustion engine according to FIG. 1 with exhaust-side catalyst and heat exchanger for the removal of thermal exhaust energy, as well as with end silencer.

[0060] Identical parts are marked with the same reference symbols in all figures.

[0061] The internal combustion engine 1, schematically depicted in FIG. 1, is designed as a two-stroke engine according to EP 4 001 609 A1. It comprises a number of cylinders 2, of which only one is shown in FIG. 1, and in each of which a working piston 4 is guided. The working piston 4 acts on a crankshaft 8 via a connecting rod 6. Depending on the design and construction of the internal combustion engine 1, the working piston(s) 4 of several or all cylinders 2 can act on a common crankshaft 8. In a conventional design, the working chamber 10 is located within the cylinder 2, in which a compressed fuel-air mixture is combusted during the power stroke of the cylinder 2. In response, the working piston 4, which is slidably arranged within the cylinder 2, executes a power stroke, thereby driving the crankshaft 8 and performing work.After completion of the working stroke, i.e. after expansion of the burned working gas in cylinder 2 and reaching the so-called "exhaust opening", the burned working gas is fed as exhaust gas during an exhaust stroke of cylinder 2 to an exhaust system 12 connected to it on the exhaust side.

[0062] For the gas exchanges required for the operation of cylinder 2, 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 the exhaust port(s) 16 opening into the working chamber 10. The gas exchanges in the working chamber 10 are controlled in the manner of a conventional two-stroke engine by the up and down movement of the working piston 4 in the cylinder, which opens or closes the gas inlet windows 14 and the gas exhaust port 16 as appropriate. In a conventional port-controlled scavenging system in a two-stroke engine, the gas exhaust port 16, and thus the exhaust port, opens first during the expansion stroke to reduce the cylinder pressure by releasing exhaust gas (necessary pre-exhaust).Subsequently, the gas intake ports 14 open, and thus the transfer ports(s) that admit fresh gas from the crankcase (or turbocharger) into cylinder 2. The type of cylinder scavenging (reverse flow, parallel flow, etc.) is irrelevant, since all methods aim to expel the exhaust gas in cylinder 2 with the incoming fresh gas from cylinder 2 with the least possible mixing of the two gas phases.

[0063] In conventional slotted two-stroke combustion engines 1, 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. To remedy this problem, the two-stroke combustion engine 1 has a switchable multi-way changeover 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 changeover 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.

[0064] Towards the end of the charge exchange, when the gas inlet ports 14 are already closed, 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 shown in FIG. 1. 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 provides boost or recharging in cylinder 2 at the end of the charge exchange.

[0065] Regarding the dimensioning of the components, the timing design of the port windows in the cylinder and the overall design of the two-stroke internal combustion engine 1 are preferably taken into account. Smaller port volumes support high specific engine outputs, while larger volumes reduce scavenging losses. The volume of the section of the exhaust port 18 located between the gas exhaust port 16 of the 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 the cylinder 2.

[0066] In the embodiment shown in the exemplary embodiment, the fresh gas pressure line 32, in a configuration considered to be independently inventive, is itself connected on 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. In the exemplary embodiment, the exhaust gas-driven charging device 40 is also configured, in a configuration considered to be independently inventive, as an exhaust gas charging pump, as known, for example, from EP 2 846 019 A1, EP 2 846 020 A1, EP 3 061 970 A1, or EP 3282 109 A1. It comprises a diaphragm pump 42, which is connected on its secondary side via the fresh gas pressure line 32 to the multi-way switching valve 30 and on its primary side to the outlet channel 18.For this purpose, 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 as required.

[0067] To illustrate the operation of the aforementioned components, the two-stroke internal combustion engine 1 according to FIG. 1 is shown in FIGS. 2 to 4 at various times during the exhaust stroke.

[0068] FIG. 2 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. In this 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 during which the exhaust port 18 is exclusively connected to the exhaust gas charging pump when it opens over the piston edge (exhaust gas charging) 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.

[0069] In the subsequent second phase, corresponding to the operating position of the two-stroke combustion engine 1 according to FIG. 3, the multi-way switching valve 30 has already opened the flow path from the exhaust port 18 to the exhaust system 12 and the flow path of the relaxed, reverse exhaust mass flow from the diaphragm pump 42 of the exhaust gas charging pump to the exhaust system 12. In this phase, the primary side of the exhaust gas 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 gas charging pump 42. This corresponds to the operating mode generally intended for the exhaust gas charging pump 42.

[0070] Figure 4 shows the two-stroke internal combustion engine 1 in a third or final phase of the charge exchange, corresponding to the situation in Figure 1. In this final phase, when the gas outlet window 16 in the 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. In addition, 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.This forces compressed or pressurized fresh gas into the exhaust port 18, which in turn pushes the gas components located in the exhaust port 18, primarily the scavenging losses from cylinder 2, back into it and also causes fresh gas recharging or boosting in cylinder 2. Advantageously, the opening of the fresh gas pressure line in the exhaust port occurs 0-30° crank angle before the closing of the fresh gas transfer ports in the cylinder.

[0071] Finally, the gas outlet window 16 is closed by the upward-moving working piston 4, thus completing the charge exchange. Naturally, several cylinders 2 can be supplied by an exhaust gas charging pump 42 via one or more such valve devices 30.

[0072] 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 exhaust port 18. In the exemplary embodiment, it is designed as a roller valve, in particular as a cylindrical, axially rotatable rotary valve 44 in a stationary control housing 46. In an advantageous embodiment, the control housing 46 can be formed in one piece with the associated cylinder 2, which is easily achievable using casting techniques. In contrast to the embodiment known from EP 4 001 609 A1, in which the rotary valve 44 rotates synchronously with the crankshaft speed of the engine, being driven indirectly by it via intermediate elements such as toothed belts, chains, or gears, or directly by mounting the rotary valve 44 directly on the crankshaft 8, or...Since the rotary valve 44 represents a part of the crankshaft 8, the present embodiment, according to one aspect of the invention, is designed for a further improvement in efficiency and a further reduction in scavenging losses. For this purpose, 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 fact that the rotary valve 44, which in the embodiment operates at twice the crankshaft speed, also assumes its gas exchange-controlling position at bottom dead center (BDC) in the top dead center (TDC) region is not detrimental, since the exhaust port 16 in the cylinder 2 is closed by the piston 4 at this moment, and thus no significant gas exchange takes place at the rotary valve 44 in this piston or crankshaft phase position.The fact that the rotary valve 44 connects the exhaust port 18 with the fresh gas pressure line 32 in the TDC range of the piston 4 is not considered to be essentially disadvantageous, since the volume in the exhaust port 18 between cylinder 2 and rotary valve 44 can be considered very small compared to the stroke volume and also the delivery volume of the fresh gas compressor, and small amounts of fresh gas flowing into the exhaust port in the TDC range are tolerable as a welcome temperature-reducing leakage.

[0073] The control housing 46 is shown enlarged in perspective view in FIG. 5. In addition, FIGS. 6 to 8 show the associated rotary valve 44, designed as a roller valve, in perspective views (FIGS. 6 and 7) and in cross-section (FIG. 8).

[0074] The control housing 46 shown in FIG. 5, which is advantageously and according to one aspect of the invention is designed 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 rotary valve 44, which is shaped like a roller or cylinder, can be inserted. Together, the rotary valve 44 and the control housing 46 thus combined form the multi-way switching valve 30. As can be seen in FIG. 5, the control housing 46 has 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 gas compressor, and a channel 57 that is connected to the outlet channel 18.

[0075] The channel elements 54, which open 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. As can be clearly seen in the illustrations of the rotary valve 44 in FIGS. 6 to 8, the channel segments 56 in the rotary valve 44 are responsible for the fresh gas flow, with a further channel segment 58 being provided for the exhaust gas flow. The base body of the rotary valve 44 forms the control edges 60. These define unambiguous control times according to which the rotary valve 44 passes over the corresponding channels 50, 52, 54, 57 in the control housing 46 in order to carry out the gas exchange in accordance with the concept of the invention.Thus, in the exemplary embodiment, the rotary valve 44 is provided with control cutouts that correspond to channel windows in the control housing 46.

[0076] The aforementioned doubling of the rotational speed of the rotary valve 44 relative to the crankshaft speed, as provided for in one aspect of the invention, 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. In particular, according to an aspect considered to be independently inventive, 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. In principle, however, it is possible (in particular by suitable specification of the respective tangential segment length of the respective track 61a,b and, if necessary,further) a desired switching function, which also includes the synchronous control of the gas flows even at the increased speed of the rotary valve 44 as explained above, is produced by the suitable arrangement of the functionally different channel routing of several tracks 61a,b.

[0077] A suitable design of the channel guides in the body of the rotary valve 44 itself, by appropriate contouring of the channel segments 56 and 58, in turn enables the confluence of, or controls, the channels of different tracks 61 in the control housing 46. In the illustration in FIG. 5, the tracks 61a,b in the control housing 46 relate to the opening cross-sections for the compressed fresh gas to be backflushed via the outlet channel 18 into the cylinder 2 and correspond at the rotary valve 44 to the channel segments (56a and 56b) designed as channel pockets, which are brought together in the region 56c to push compressed fresh gas via the channel 62 in the control housing 46 into the outlet channel 18. This concept of appropriately designed tracks 61 is considered to be independent and autonomously inventive.

[0078] FIG. 9 shows a longitudinal section of the assembled multi-way diverter valve 30, i.e., with the rotary valve 44 inserted into the receiving channel 48. The channel segments 56, 58 of the rotary valve 44 seal against the channel elements 50, 52, 54, 57 in the control housing 46 preferably and according to an independent inventive aspect, without contact via a gap 64, which is in the range of 3 to 25 hundredths of a millimeter. According to aspects of the invention, 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. For this purpose, according to aspects of the invention, for example, sliding varnishes or plastics can be provided as soft coatings, which are partially worn away during the break-in phase without causing consequential damage and thus create a minimal sealing gap between the rotary valve 44 and the control housing 46.Preferably, 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.

[0079] Furthermore, 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. Advantageously, the cooling of the rotating rotary valve 44, as can also be clearly seen in FIG. 9, is implemented with a coaxial inlet and outlet (66, 68) of coolant in counterflow 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.

[0080] 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 shown in FIG. 9. To protect the rolling bearings 78 and 58, respectively,To protect radial shaft seals from excessively high gas pressures from the rotary valve control area, sliding seals 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 seal 80 and the rolling bearing 78 is advantageously fed to the engine's intake air via bypass channels.

[0081] The intended use of the engine concept described herein as a drive source for power generators, heat pumps, or vehicles, considered particularly advantageous and independently inventive, 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. To achieve these last three requirements in particular, 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. This is illustrated in the conceptual representation of the independently inventive concept shown in FIG.Since 10 is removable, 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.

[0082] Further aspects, each of which, individually or in any combination, is considered to be independently inventive, are explained in more detail below. FIGS. 10a and 10b schematically show the two-stroke internal combustion engine 1 in a vertical section with a separate oil bath 82 for the crankshaft main bearings 81 and the drive gears 84, with a separate supply 86 of loss oil through one of the crankshaft journals to the crankpin 88 of the connecting rod bearing 90, with a 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 and for the upper connecting rod bearing 100. FIG. 10b shows a vertical section perpendicular to 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 blow-off 112-120 for piston pin lubrication. FIG.Figure 10c, on the other hand, 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.

[0083] FIG. 10a shows, in particular in a 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. In this design, the crankshaft main bearings 81 no longer require lubrication via a loss oil supply. The connecting rod bearing 90 of the crankshaft crankpin 88 is supplied directly with metered loss oil via channels machined into the crankshaft 8, preferably, as shown in FIG. 10, via an axial inlet bore 104 in one of the crankshaft crank discs, the loss oil being conveyed via bores in the crankshaft crank disc to the crankshaft crankpin 88 and one or more outlet bores 106 directly at the connecting rod bearing.

[0084] 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 supply of lubricating oil to the piston pin bearing 88, including improved cooling of the piston crown, 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 shown in FIG. 10b.

[0085] Another embodiment for lubricating the piston pin bearing 88, considered to be independently inventive, provides for an oil-air supply to the piston pin 98 and its bearing in the connecting rod 6. Figures 10a and 10b show 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.

[0086] 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. According to one aspect of the invention, metered oil can be added to the pressure line 118 near the cylinder wall via a supply line 122.

[0087] 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 is reached in the crankcase 110. Up to this point, fresh gas, possibly containing oil, is pumped from the crankcase 110 via the check valve 114 to the pressure accumulator 116, where it remains. This is because the piston skirt closes the inlet of the pressure line 118 into the working chamber 10 at the inlet point 120, thus preventing a short-circuit flow inside the crankcase 110.

[0088] As the piston 4 continues its movement towards bottom dead center after the scavenging ports 14 have opened, the fresh gas pressure in the crankcase 110 is reduced by the flow of fresh gas into the working chamber 10 of the cylinder 2. However, the maximum stored pressure of the fresh gas mass in the pressure accumulator 116 remains approximately constant. Near the bottom dead center of the piston 4, the radial bore 96 in the piston skirt, clearly visible in FIG. 10c, overlaps the bore of the pressure line 118 in the cylinder wall, which forms the inlet 120. The bore 96 in the piston skirt opens inside the piston 4 near the bearing of the piston pin 98 and preferably points in its direction.

[0089] Due to the decreasing pressure in the crankcase 110 during cylinder scavenging and the opening of the pressure line 118 through the radial bore 96 in the piston 4, fresh gas now flows from the pressure accumulator 116, carrying with it previously introduced oil loss from the pressure line 118, through the bore 96 in the piston 4 to the piston pin 98 and its bearing. The fresh gas circulation of this system is self-regulating, with the amount of oil supplied being freely selectable. This operating concept is considered to be independently inventive.

[0090] Advantageously, fresh gas-oil injection can be applied to the bearing of the piston pin 98 from both sides, as shown in FIG. 10a with reference to the two injection points 120 and in FIG. 10c with reference to the two radial bores 96. The pressure accumulator 116, including its lines, preferably has a volume of 2-10% of the stroke volume of the working cylinder 2.

[0091] FIG. 10b shows the supply of oil to the cylinder wall through corresponding channels and lubrication pockets in the cylinder bore below the exhaust port 18 and similarly on the opposite side, 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. Preferably, the oil supply is located below the exhaust ports or on the opposite side of the cylinder bore, since the piston skirt bears against it in these areas due to the pivoting movement of the connecting rod under load.

[0092] According to an aspect considered to be independently inventive, 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. According to aspects of the invention, this can be achieved by adding liquid water to the gas mixture in the working cylinder or directly to the fuel. Particularly in the case of a turbocharged engine 2, 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.

[0093] According to aspects of the invention, such a continuous supply of water to the motor 1, considered advantageous in stationary applications such as charging stations for electric vehicles, power generators, or heat pump drives, can be ensured either via existing water supply networks near buildings or by condensing water vapor from the engine exhaust gases. In addition to self-contained water injection independent of the fuel used, another aspect of the invention offers the very simple solution of adding water to the fuel in a desired ratio, which can be achieved with water-miscible fuels such as methanol, ethanol, and ammonia.

[0094] FIG. 11 shows a longitudinal section of a cylinder 2 of the internal combustion engine 1 with such a water injection system, which is considered to be independently inventive. For this purpose, 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. In this position 184c, the water injection valve is permanently exposed to only comparatively low pressures and temperatures, which, according to one aspect of the invention, enables a simple and cost-effective design.

[0095] Water injection during the scavenging phase, with the onset of evaporation, effectively 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. Both the lower gas temperature of the fuel-air mixture at the start of combustion and the water vapor content, with its considerable heat capacity, significantly reduce the peak temperature during combustion, thus contributing to a substantial reduction in nitrogen oxide formation. The temperature reduction during the compression phase generally shifts the knock limit to higher compression ratios, which can be used to increase engine efficiency.

[0096] According to one aspect of the invention, the indicated mean effective pressure can be significantly increased via the piston movement by the amount of water introduced into cylinder 2. For this purpose, according to a separate inventive aspect, the 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 pressure increase associated with the evaporation thus directly benefits the indicated piston work.

[0097] According to aspects of the invention, 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. This design allows for free selection of the injection timing, with the option of advancing the injection timing even before the ignition timing in order to achieve a compromise between high expansion work and reduction of nitrogen oxide formation. While accommodating a water injection nozzle 185 in the cylinder head 186 does not pose any problems in terms of installation space, this design is the most expensive, as the injection nozzle 185 must be operated under high temperatures and water pressures.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. However, 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.

[0098] Furthermore, 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. This not only increases the thermodynamic efficiency of engine 1 but also reduces emissions such as CO, HC, and NOx. Particularly in the two-stroke engine 1 under partial load with its high residual exhaust gas contents in the combustion chamber 10, compression ignition enables reliable ignition at low loads and with lean fuel-air mixtures. FIG.Figure 12 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 seals the combustion chamber 195 against a fixed squish insert 196 with a squish gap 198 via sealing rings 194. Regardless of the axial position of the combustion chamber insert 192, 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.

[0099] According to one aspect of the invention, 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 "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 working gas is reliably reached before the thermodynamically ideal ignition point.The resulting premature ignition timing, caused by the excessively high compression end temperature, is now regulated to the desired ignition timing by the individual addition (injection) of water for each working stroke, with its temperature-reducing evaporative cooling. Further parameters for controlling HCCI combustion can include combustion chamber temperature and crankshaft unevenness, the measured values ​​of which can be detected by sensors.

[0100] Based on these considerations, and in order to particularly promote the desired low maintenance and durability of the engine, 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. For this purpose, according to a separate aspect of the invention, 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 81 (see also FIG. 10a). The bearing 81 is thus hermetically and tribologically separated from the interior of the crankcase 110 and also does not require a lubricant supply from inside the crankcase.Advantageously, all crankshaft main bearings 81 are separated from the respective interiors of the corresponding crankcases 110. Advantageously, 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. However, alternatively, 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 can be seen, for example, in the illustrations in FIGS. 13 and 14. These show the respective gear cascade, which distributes the oil from the oil bath 82 through the meshing of several gears 216 in the enclosed space of the gear drive.

[0101] 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. In such an embodiment, 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.

[0102] Advantageously, and according to one aspect of the invention, the multi-way switching valve is driven directly or indirectly by the crankshaft 8. To achieve, in a particularly simple manner, the operation of 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. In an embodiment considered to be independently inventive, the transmission ratio of the transmission 210 is 1 : n, where n is an integer natural number. Preferably, n = 2, so that the multi-way switching valve 30 operates at twice the speed of the crankshaft 8.

[0103] To meet the high demands for the required NVH (noise, vibration, and harshness) behavior in applications close to buildings, the engine 1 should operate completely or as far as possible without vibration, since free forces or moments at the engine could be transmitted through its foundation into the building or the ground. According to one aspect of the invention, it is therefore provided that, for largely force- and moment-free operation, the two-stroke internal combustion engine 1 is designed as a two-cylinder engine with the features of a "true" boxer engine 212. Such a boxer engine 212 is shown in frontal view in FIG. 13. The two cylinder axes 214 are arranged coaxially and in opposite directions to each other. Two separate crankshafts 8 arranged parallel to each other (i.e., one for each cylinder 2) are provided, the axes of rotation of which intersect the common cylinder axis 214, and which are advantageously coupled to each other at the same speed and in opposite directions via a gear pair 216.

[0104] The pistons of this boxer engine 212 move in opposite directions during operation. With this design, no free mass moments occur, and the free forces of the first and second order cancel each other out due to the counter-rotating pistons. The mechanical coupling of the crankshafts 8 and rotary valves 44 is advantageously achieved, and according to an aspect of the invention, as shown in FIG. 13a, via spur gears 218, each of which is assigned to one of the rotary valves 44, and which are each coupled to the gear pair 216 via an intermediate gear 220 to form the transmission 210. By suitable selection of geometry, in particular a suitable selection of the number of teeth of the respective gears, the transmission ratio of the transmission 210 is set to 1:2, according to the design in the exemplary embodiment.

[0105] An alternative design for a two-cylinder engine, also considered advantageous and independently inventive, but exhibiting (comparatively low) free inertial forces of the second order, is the arrangement of the two cylinder units (cylinders with crankshaft assembly) as a so-called tandem engine 230, as shown in FIG. 14. This engine has two axially parallel, counter-rotating crankshafts 8 of the same speed, which are advantageously coupled – analogous to the above-described embodiment of the boxer engine 212 – as shown in FIG. 14a, 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. In this embodiment, the pistons move synchronously "up and down" in their respective cylinders 2.In this embodiment, the free second-order inertial forces can be balanced, according to an aspect considered independently inventive, by the two counter-rotating rotary valves 44 and their twice the crankshaft speed, provided these have a corresponding imbalance and act as a Lancaster differential. In principle, in the case of two coupled, axially parallel, and counter-rotating crankshafts running at the same speed, a mass balance of the free first- and second-order forces can always be achieved in conjunction with a Lancaster differential, regardless of the cylinder bank angle. A prerequisite for this is that the two crank mechanisms (pistons, connecting rods, crankshaft) are geometrically identical, so that the cylinders operate synchronously, i.e., the pistons reach their dead centers at the same time and the cylinders lie in a plane perpendicular to the crankshaft axis.The boxer engine (180° cylinder bank angle) and tandem engine (0° cylinder bank angle) configurations represent only the extremes of the cylinder bank angle between 0 and 180°. Balancing the free second-order forces with rotating counterweights via a Lancaster balancer is unnecessary at a 180° cylinder bank angle and then increases gradually with increasing unbalanced mass in the Lancaster balancer up to a cylinder bank angle of 0° (tandem engine). Cylinder bank angles smaller than 90°, especially 10–45°, are considered particularly advantageous, as these geometries allow for the construction of especially compact (V-)engines that require minimal space. Preferably, the design of such a compact V-engine is characterized by two vertically split crankcase halves onto which two separate cylinder units are mounted.The cylinder units are preferably manufactured as a single unit with the rotary valve housing. Preferably, the exhaust ports of the cylinders with the rotary valve point outwards in the direction of the crankshaft axes. Even with narrow cylinder bank angles (10-30°), there is sufficient space between the cylinders, preferably to integrate an electric starter motor and also to install fuel injectors on the cylinders, which preferably inject fuel into the respective cylinder opposite the exhaust port system (see image). The starter motor can act on a gear connected to a crankshaft stub or on a gear directly connected to a consumer such as a generator, compressor, etc. The cylinder head, the cylinders together with their rotary valve housing, and the crankcase are preferably liquid-cooled (see image).To achieve a torsionally rigid crankshaft assembly despite the multi-part construction of the crankshaft, which is a technical necessity due to the roller bearing of the undivided connecting rod foot, the torque-transmitting crankshaft ends, which also carry the spur gears for the torsionally rigid coupling of the two crankshafts, can be supported by an additional support bearing (see image). This support bearing is preferably located in the crankcase cover, which seals the gear drive and its oil bath from the outside.

[0106] Deviations from the ideal cylinder arrangements of the "true" boxer or tandem engine 212, 230, or even V-engine described here could cause vibrations due to free forces or moments, which, up to a certain degree, preferably around 15%, may be considered acceptable in practice. In both the boxer and tandem arrangements of two cylinder units, both cylinders 2 operate in phase, meaning they also fire simultaneously. Preferably, the two crankcases are hermetically connected and equipped with a common throttle valve for fresh air supply and control. This ensures a uniform delivery rate in the common crankcase pump and avoids synchronization problems associated 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, for example, to form four-, six-, or eight-cylinder engines. Components of a metallic engine block are advantageously manufactured as castings in series production, which then undergo final machining. Particularly advantageous design features are provided for the engine designs discussed here, the "true boxer engine 212" (FIG. 13) and the "tandem engine 230" (FIG. 14), with regard to the castability of individual components, ease of assembly, and reduction of the number of components. The true boxer engine 212 according to FIG. 13b is preferably formed from a cylinder base 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 both cylinders 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. In the case of crankcase pumps located inside the common crankcase, the throttle valve on the cylinder base housing 234 supplies the engine. The advantageous design of a tandem engine 230 according to FIG. 14b consists of a one-piece cylinder housing 239 for both cylinders 2, including the housing bosses for the rotary valves 30, a cylinder base housing 234 which is bolted to the cylinder housing 239 via a parting plane 236 that intersects both crankshaft main bearings 81 in the center, and a cylinder head 186 mounted on the cylinder housing 239.In the tandem engine 230, both crankcases inside the crankcase 110 are also connected and supplied with fresh air via a common throttle valve.

[0107] In contrast, considering a single-cylinder engine of the aforementioned two-stroke design, 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. To reduce the overall size, 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.Thus, the 1:2 speed ratio of the respective crankshaft 8 to the rotary valve 44, as provided for in one aspect of the invention, can be adjusted, according to an independently inventive aspect, by means of a drive of the respective rotary valve 44 directly 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. According to another aspect of the invention, the twice the 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 or a V-engine, in which, as shown in FIG. 14, two counter-rotating, opposing rotary valves 44, one for each cylinder 2, are provided.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.

[0108] The lubrication of the spur gears 218 is advantageously carried out, as shown, from the oil bath 82 shared with the crankshaft main bearings 81, advantageously 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, the tandem engine 230, or the V-engine described above, 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. To reduce noise, the spur gears 218 can also be helical.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 via 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 atmosphere.

[0109] Due to the design-related absence of combustion gases in the oil sump and the prevailing low temperatures there, 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.

[0110] In all variants of the two-stroke internal combustion engine 1 described above, 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.

[0111] Some areas of application for the two-stroke internal combustion engine 1 of the above-described design, optionally with some or all of the additionally described embodiments, are considered particularly preferred and the corresponding use of the two-stroke internal combustion engine 1 is considered to be independently inventive.

[0112] For example, FIGS. 15, 16, and 17 each show, in the manner of a possible and preferred stationary use of the two-stroke combustion engine 1, a generator unit 240, also referred to as a "combined power plant," which is considered to have an independent inventive design. The generator unit 240 comprises such a two-stroke combustion engine 1, which is connected via a drive shaft 242 to a power generator 244 and drives the latter during operation. FIG. 15 shows an embodiment considered to have an independent inventive design in which the power generator 244 is electrically connected on its output side to a wallbox 246 for charging an electric vehicle 248 near a building 250. In contrast, in the embodiment shown in FIG. 16, which is also considered to have an independent inventive design, the power generator 244 of the combined power plant 240 is electrically connected on its output side to a heat pump 252 installed in the building 250 for heating purposes.Naturally, in an embodiment also considered independently inventive, these two aspects can also be combined; that is, the power generator 244 can be connected on the output side to both the wall-box 246 and the heat pump 252, with the output line being suitably controlled and distributed to these components as needed. Preferably, the power generator 244 is also designed for particularly favorable NVH (noise, vibration, and harshness) characteristics to avoid disturbing the residents' sleep. When connected to the wall-box 246, an electrical output of approximately 5–25 kW is preferably provided, depending on the size and number of vehicles to be charged. The waste heat from the combustion engine 1 can advantageously be supplied to the surrounding buildings 250 for heating or hot water preparation (combined heat and power).In addition to using the power generator as a charging station for an electric vehicle 248, it can also serve as a grid-supporting or grid-independent power source for driving the heat pump 252.

[0113] When operating an electrically driven heat pump 252 independently of the grid, 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. In conjunction with the heat pump 252, 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, advantageously to a temperature level below the condensation temperature of the water content.This use of engine waste heat increases the inlet temperature of the working fluid before the compressor 270, thereby raising its COP (Coefficient of Performance) by reducing the required temperature lift at a given heating supply temperature. It makes no difference whether the evaporator 255 draws its heat from the ambient air or the ground.

[0114] As can also be seen from FIGS. 15 and 16, in both cases, 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.

[0115] In such a design, particular attention is paid to the soundproofing of the motor 1 in conjunction with the electric generator 244, which, preferably located near buildings, should also be able to run overnight. 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 material can advantageously be made of multiple layers of materials with different sound absorption coefficients. A horizontal division plane 256 of the soundproof shell 254, forming a removable upper cover 258, is advantageously provided, allowing for particularly easy maintenance of the internally arranged combined heat and power plant 240 and its components.

[0116] 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.

[0117] According to a further preferred embodiment, the complete combined heat and 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. According to aspects of the invention, 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.

[0118] 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, the springs being attached to the housing 260 or to a subframe. 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 certainly offers aesthetic advantages due to its concealment. The arrangement shown in FIG.Figure 17 also shows the concept of a heat pump 252, whose compressor 270 can be driven by a two-stroke combustion engine 1 of the type described. In the embodiment shown in Figure 16, this drive is implemented as an indirect drive, namely via the generator 244. Alternatively, and in a manner considered to be independently inventive, such a drive can also be provided directly, as can be seen from the illustration in Figure 17.

[0119] In such a case of direct drive of the compressor 270 by the motor 1, 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. Alternatively, and in an embodiment also considered independently inventive, the compressor 270 of the heat pump 252 and the power generator 244 can both be provided together in the underground arrangement, preferably being operated alternately, for example by mechanical switching. In particular, in such a synergistic combination, considered independently inventive, 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.

[0120] In contrast to the aspects explained above, FIG. 18, for example, shows a vehicle 280 with such a two-stroke internal combustion engine 1, which is considered to be independently inventive, as a possible and preferred mobile application of the two-stroke internal combustion engine 1. The vehicle 280 is shown in FIG. 18a as an exemplary perspective top view of the chassis 282 of a four-wheeled passenger vehicle; naturally, the construction concept described below is also applicable analogously to two-wheelers or commercial vehicles. FIG. 18b shows the chassis 282 of the vehicle 280 in a side view, FIG. 18c in a front view, and FIG. 18d in a top view.

[0121] As can be seen from the illustrations in FIG. 18, in the exemplary embodiment, the two-stroke internal combustion engine 1, in this case designed 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. In the exemplary embodiment, 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.

[0122] According to one aspect of the invention, 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. Furthermore, and in addition to these components, 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.

[0123] The combination, as provided for in this aspect of the invention, of the comparatively inexpensive two-stroke engine 1 for on-board power generation, in conjunction with a comparatively 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. In principle, 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.Similarly, the placement of the two-stroke combustion engine 1 near the rear axle 284 is not mandatory, but can generally be freely chosen within the vehicle 280. Furthermore, engine designs with tandem cylinder arrangements are also feasible in the vehicle, as are inline engines and V-engines.

[0124] According to aspects of the invention and in a concept considered inventive independently of the foregoing, the generation and provision 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 splitting is either provided 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 provision is shown schematically in FIG.Figure 19 shows two possibilities for introducing hydrogen into the working space 10 of the cylinder 2.

[0125] The two-stroke combustion engine 1 is shown in the context 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 unit 302. This unit, in turn, is electrically connected on the inlet side to the symbolically depicted vehicle electrical system 314 280 and, on the media-side, is connected via lines (not shown) to the water tank 298 or an alternative water supply for the supply of the water to be split. The metering valve 312 allows for the sequential or continuous supply of H2 or HHO into the transfer 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 into 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. The control can be achieved via a metering valve 316 or directly via the power supply to the electrolysis unit 302.

[0126] Figure 20 schematically shows the embodiment of the two-stroke combustion engine 1, considered to be independently inventive, with an exhaust-side catalyst 320 and heat exchanger 322 for extracting thermal exhaust energy, as well as with a muffler 12. Reference numerals

[0127] 1 Two-stroke internal combustion engine

[0128] 2 cylinders

[0129] 4 working pistons

[0130] 6 connecting rods

[0131] 8 Crankshaft

[0132] 10 workroom

[0133] 12 Exhaust system

[0134] 14 gas inlet windows

[0135] 16 gas outlet windows

[0136] 18 Outlet channel

[0137] 30 multi-way diverter valve

[0138] 32 Fresh gas overpressure line

[0139] 40 Exhaust gas-operated charging device

[0140] 42 Diaphragm pump

[0141] 44 Rotary valve

[0142] 46 Control housings

[0143] 48 recording channels

[0144] 50 Outlet breakthrough

[0145] 52-channel

[0146] 54 Channel

[0147] 56 channel segment

[0148] Channel 57

[0149] 58 channel segment

[0150] 60 Control edge

[0151] 61 tracks

[0152] 62-channel

[0153] 64 columns

[0154] 65 coating

[0155] 66 Inflow

[0156] 68 Drain

[0157] 70 Rotation axis

[0158] 72 Seal

[0159] 74 Shaft end rolling bearing sliding seal ring crankshaft main bearing oil bath

[0160] Gear feed, crankshaft journal, connecting rod bearing, oil loss feed, bore, piston pin, upper connecting rod bearing, shaft seal, inlet bore, outlet bore, crankcase, pressure line, check valve, pressure accumulator, pressure line, feed point, supply line, position, water injection valve, cylinder head, shot channel, two-stroke cylinder head, combustion chamber insert, combustion chamber sealing ring

[0161] Squish clearance insert, squish clearance, transmission gearbox, boxer engine, cylinder shaft, gear pair, spur gear, intermediate gear, tandem engine, gear, cylinder base housing, parting plane, cylinder block, cylinder housing, generator block, drive axle, power generator, wall box, electric vehicle, building, heat pump, heat exchanger, soundproofing shell, evaporator, parting plane, cover, tray, line, pipe, roofing, compressor, electric motor, vehicle, chassis, rear axle, battery pack

[0162] Sodium-ion battery, drive motor, differential gear, front axle, fuel tank, water tank, hot water tank, electrolysis system, drive module, metering valve, on-board electrical system, metering valve, feed-in position

Claims

Claims 1. Two-stroke internal combustion engine (1) with a number of cylinders (2) in which pressurized working gas can expand to perform work, thereby driving a crankshaft (8), wherein the gas outlet port (16) of the or each cylinder (2) is connected to an exhaust system (12) via an exhaust port (18), and wherein a switchable multi-way switching valve (30) is arranged in the exhaust port (18), via which the gas outlet port (16) is connected to both the exhaust system (12) and a fresh gas pressure line (32), characterized in that the multi-way switching valve (30) is designed as a controllable rotary valve (44) for operation at a speed that is an integer multiple of the speed of the crankshaft (8) by a factor of at least two.

2. Two-stroke internal combustion engine (1) according to claim 1, the multi-way switching valve (30) of which is driven by the crankshaft (8).

3. Two-stroke internal combustion engine (1) according to claim 2, the multi-way switching valve (30) of which is connected on the drive side to the crankshaft (8) via a transmission transmission (210) with a transmission ratio 1 :n, wherein n is an integer natural number.

4. Two-stroke internal combustion engine (1) according to claim 3, the transmission transmission (210) of which is designed as a gear transmission, the gears (216) of which are connected to a common oil circuit for the supply of lubricant.

5. Two-stroke internal combustion engine (1) according to one of claims 1 to 4, the multi-way switching valve (30) of which is designed for operation at twice the speed of the crankshaft (8).

6. Two-stroke internal combustion engine (1) according to one of claims 1 to 5, which is designed as a two-cylinder engine, wherein each of the two cylinders (2) is assigned a crankshaft (8), and wherein the crankshafts (8) rotate in opposite directions to each other when the engine (1) is operated and are coupled to each other via a number of gears (216).

7. Generator block (240) comprising a two-stroke internal combustion engine (1) according to one of claims 1 to 6, and comprising a current generator (244) driven by the two-stroke internal combustion engine (1).

8. Generator block (240) according to claim 7, wherein both the two-stroke combustion engine (1) and the power generator (244) are installed underground.

9. Vehicle (280) with a two-stroke internal combustion engine (1) according to one of claims 1 to 6.

10. Vehicle (280) according to claim 9, further comprising a current generator (244) driven by the two-stroke combustion engine (1), which in turn is connected on its output side to a storage battery (286), preferably to a sodium-ion battery (288) provided as an energy storage device.

11. Heat pump (252) with a compressor (270) which can be driven by a two-stroke internal combustion engine (1) according to one of claims 1 to 6.

12. Internal combustion engine according to claim 6, wherein two working cylinders are combined to form a two-cylinder V-engine, characterized in that the two working cylinders are arranged in a V-shape with a bank angle between 0° and 180°, preferably with a bank angle of 10°-45°, with pistons moving in the same direction, wherein each piston acts on its own crankshaft, wherein the crankshafts rotate in opposite directions and are coupled to each other via gears.

13. Internal combustion engine according to claims 6 and 12, wherein the free second-order inertial forces are compensated by a Lancaster mass balance, characterized in that the required rotating unbalance masses are arranged on the two rotary valves of a cylinder pair which operate at twice the crankshaft speed.

Images