METHOD FOR OPERATING A TWO-STROKE INTERNAL COMBUSTION ENGINE AND TWO-STROKE INTERNAL COMBUSTION ENGINE

Description

[0001] The invention relates to a method for operating a two-stroke internal combustion engine having a number of cylinders in which pressurized working gas can expand to perform work, wherein the gas outlet port of the or each cylinder is connected to an exhaust system via an exhaust port. It further relates to such a two-stroke internal combustion engine for implementing the method.

[0002] Internal combustion engines or heat engines are used in a wide variety of applications, for example, as drive units for motor vehicles or in stationary versions in industrial plants or other technical facilities. An internal combustion engine or heat engine can be operated using a gasoline or diesel engine, using various liquid or gaseous fuels. What the various designs have in common is that a compressed fuel-air mixture is usually combusted in a cylinder's working chamber (commonly referred to as the "combustion chamber") (or, analogously, in a rotary disk chamber in a Wankel engine). In response, a piston movably arranged in the respective cylinder executes a power stroke, thereby driving an engine shaft to produce power.After the completion of the working stroke, i.e. after expansion of the burned working gas in the cylinder, the burned working gas is fed as exhaust gas during an exhaust stroke of the respective cylinder to an exhaust system connected to the exhaust side of the cylinder.

[0003] In a common design, the internal combustion engine can be a 4-stroke engine, in which the exhaust system leading to the exhaust system is released via an exhaust valve located in the cylinder. In an alternative common design, the internal combustion engine can also be a 2-stroke engine, in which the exhaust system leading to the exhaust system is released in a slot-controlled manner, with the piston sliding over exhaust slots located in the cylinder in the final phase of its expansion stroke and then releasing them. Likewise, a corresponding intake valve system is usually provided on the gas inlet side in 4-stroke engines, whereas in 2-stroke engines this functionality is also provided by scavenging or transfer slots located in the cylinder jacket. There are also valve systems based on rotary valves, in which channel cross-sections in the cylinder head are opened or closed by rotating rotary valves.Rotary valve control systems are used for two- and four-stroke cycles. Therefore, the terms "intake valve system" and "exhaust valve system" used here also explicitly refer, in the case of two-stroke engines, to the corresponding slot arrangements in the cylinder barrel, through which the corresponding valve functions are implemented, analogous to their actual embodiment in four-stroke engines using poppet valves or rotary valves.

[0004] The most common design in piston engines today is the four-stroke engine, regardless of the combustion process (diesel or gasoline). Engines with a displacement of approximately one liter or more are largely equipped with exhaust gas turbocharging. The four-stroke engine dominates both the entire passenger car engine sector, including commercial vehicle drives, and large engines. In low-powered passenger cars, as well as in all two-wheeled vehicles, four-stroke engines are usually naturally aspirated without turbocharging. There are three primary reasons why the four-stroke engine, in various designs, with different cylinder counts and displacements, has established itself across such a wide range of applications.

[0005] Firstly, the extensive separation of the charge cycle between fresh and exhaust gases is important to note, which avoids direct scavenging losses of unburned fresh gas into the exhaust tract and thus fundamentally enables low-emission drive systems. In this context and in subsequent use, "fresh gas" refers to both pure air and air-fuel mixtures, depending on how the mixture is prepared locally and temporarily. Secondly, in conjunction with exhaust gas turbocharging, the overall efficiency of the engine can be increased. Due to the increased charge density, friction and wall heat losses hardly increase, while the effective mean effective pressure of the engine is increased roughly proportionally to the increase in air density. The utilization of the otherwise unused exhaust gas enthalpy by the turbocharger, in conjunction with a reduction in engine speed (so-called "downspeeding") orReducing the displacement (cylinder) ("downsizing") allows the engine to operate at a thermodynamically more favorable operating point with a higher volumetric efficiency and further de-throttling. Third, the four-stroke gasoline engine can operate very well at part-load compared to a two-stroke engine by expelling residual gas from the cylinder. The high residual gas content in a two-stroke gasoline engine at part-load operation, on the other hand, can cause misfiring or combustion failure, which is associated with very high HC or CO emissions.

[0006] As a result, the two-stroke engine is currently only used industrially in a number of completely different niche applications. One such example is large engines for marine propulsion, which operate using the two-stroke diesel process with exhaust gas turbocharging. The fresh gas is scavenged using slot control (transfer ports at the bottom dead center) in the cylinder and exhaust valves in the cylinder head (head-controlled) as a direct-flow scavenging system. The engines are often designed for extremely long strokes, thanks to a crosshead connecting rod linkage. The diesel combustion process eliminates fuel scavenging losses, and it also allows for trouble-free part-load operation, as the exhaust gas turbocharging (positive scavenging gradient) adequately scavenges exhaust gas from the cylinder even at part-load.The asymmetric timing diagram of the exhaust valves allows the exhaust end to be positioned ahead of the scavenging end of the transfer ports, thus boosting the engine. The power density, uniformity of crankshaft torque, and, above all, specific fuel consumption are superior to all four-stroke alternatives with this two-stroke engine concept.

[0007] In addition to the inherently high-emissions nature of bunker fuel, a disadvantage is the lubricating oil consumption at the cylinder wall, which is caused by lubricant being carried along by the fresh gas at the scavenging ports and contributes to higher CO emissions. Furthermore, the design principle of the head-valved two-stroke engine is not transferable to space-optimized vehicle engines, as the crosshead crankshaft and the valve train in the cylinder head result in an unacceptable overall height of the engine. Furthermore, the thermodynamically favorable long stroke combined with the small piston diameter prevents sufficient scavenging cross-sections in the cylinder head for the operating speeds typical in passenger car engines.

[0008] Two-stroke engines with slot scavenging and supercharging are known from patent CH162817, whereby in a final phase of the exhaust stroke of the cylinder the gas-side connection between the gas outlet window and the exhaust system is interrupted and instead the gas outlet window is supplied with fresh gas under excess pressure via the exhaust channel.

[0009] Other existing industrial applications of two-stroke engines can be found in two-wheelers, especially in the 50cc class, boat drives (outboards), and portable handheld devices (chainsaws, trimmers, etc.). These are exclusively crankcase-scavenged engines with reverse scavenging and pure slot-controlled gas exchange in the cylinder. In addition to low production costs, high power density and torque uniformity are crucial for handheld devices.

[0010] In the low-priced vehicle segment, the advantages of the two-stroke engine lie in its lower production costs compared to the four-stroke design, as well as in the noticeably lower friction losses (no valve train!) and the improved torque curves at the crankshaft, which ensure noticeably improved performance in this power class. However, although vehicle manufacturers are increasingly equipping their two-stroke engines with fuel injection and have been able to significantly reduce fuel scavenging losses, the results are hardly sufficient to comply with the Euro 5 or Euro 6 emission limits in the future. The reason for this still lies in the open gas exchange of the symmetrical timing diagram of a slot-controlled two-stroke cylinder, whose exhaust port remains open up to 35° CA (CA = crankshaft) after the transfer ports close.Even with direct injection in the cylinder head, the start of injection must begin early for good homogenization of the mixture, i.e. when the exhaust port is still open, so that scavenging losses of unburned fuel in the exhaust tract are unavoidable, which lead to undesirable HC emissions.

[0011] The problem of scavenging losses of air or air-fuel mixtures described here is technically referred to as insufficient "capture efficiency" during the engine's gas cycle. The capture efficiency indicates the proportion of fresh air remaining in the working cylinder compared to the amount of fresh air scavenging actually available for combustion in the cylinder; in practice, this rarely reaches values ​​higher than 75-85%. In the two-wheeler sector, a so-called resonance exhaust system is often used to increase the volumetric efficiency, as well as to increase the capture efficiency of the two-stroke engine. The effect of this system is to generate a negative pressure in the cylinder (increased volumetric efficiency) through suitable cross-sectional changes in the exhaust pipe during the scavenging phase and to generate an overpressure in the exhaust at the end of the scavenging phase in order to push fresh gas that has entered the exhaust tract back into the cylinder (increased capture efficiency).Geometrically, such an exhaust has the shape of a "pear" with a continuously expanding exhaust pipe (diffuser) near the cylinder to generate backpressure waves early on, and a narrowing exhaust end (countercone) to reflect excess pressure waves. Following the performance-determining part of the resonance exhaust, the muffler system usually follows.

[0012] This type of exhaust system supercharges the engine via resonant vibrations of the exhaust column, and it was widely used in two-stroke engine-dominated racing of earlier years. The functionality of this resonance exhaust system essentially requires the symmetrical timing diagram of a slot-controlled engine. The main disadvantages of the resonance exhaust system are the fact that the positive resonance effects (supercharging) only work within a narrow speed range (approx. 1500 engine revolutions per minute), which is insufficient for everyday engines. Furthermore, these exhaust systems require enormous installation space, as their length is determined by the required wavelength of the resonance speed and increases with decreasing engine operating speed.In practice, exhaust lengths for racing applications are in the range of one meter and for low-speed utility engines up to two meters, which would be difficult to accommodate in vehicles, coupled with the fact that a separate exhaust system is required for each engine cylinder.

[0013] Without further measures to increase the capture efficiency of the two-stroke engine, this engine principle will likely disappear from vehicle use or even in the higher power ranges of aviation or nautical applications due to emissions (HC). CO emissions (partially burned oil), which are primarily caused by loss lubrication of the cylinder bore, have been significantly reduced in recent years. This was achieved through improved surface topographies of cylinder bores, piston skirt coatings, and piston rings, as well as the use of more high-temperature-resistant ester-based lubricants, which also burn soot-free. These measures have made it possible to reduce the oil-fuel mixture ratio in test engines to as low as 1:300, enabling absolute lubricant consumption of less than 1g / kWh.

[0014] Furthermore, if the need to supply the crankshaft via loss lubrication is eliminated, which would be possible with a separate recirculating lubrication system, lubricant consumption would be reduced even further, reaching levels close to the 0.3-0.5 g / kWh achieved by a current four-stroke engine. Finally, carbon-based "dry-lubricated" piston-cylinder pairings can ensure completely oil-free operation in the cylinder's combustion chamber.

[0015] The unfavorable part-load behavior of the two-stroke engine with high internal exhaust gas recirculation rates ("EGR rates"), which lead to misfiring, can be counteracted by turbocharging and a corresponding load point increase. Furthermore, a further shift in the operating point to higher loads is possible through "cylinder deactivation" or injection and ignition in every other working cycle ("four-stroke").

[0016] The open gas exchange of the two-stroke engine, with its highly fluctuating capture efficiency, makes stoichiometric metering of fuel to air consumption virtually impossible, which is why regulated lambda 1 operation with a three-way catalyst for exhaust gas aftertreatment is hardly feasible. Two-stroke engines are therefore typically operated with unregulated oxidation catalysts and, to advantage, with excess air. This process is cost-effective to implement, but cannot aftertreat NOx emissions. These raw NOx emissions are generally lower in two-stroke engines than in four-stroke engines because two-stroke engines produce faster mixture combustion and tend to tolerate high EGR rates. However, the untreated NOx emissions are outside current limits. One way to drastically reduce NOx formation is water injection, which uses the water's evaporation enthalpy to reduce temperature peaks during combustion.In particular, the combination of water injection combined with a carbon-based dry lubrication system for the piston and cylinder wall is considered advantageous because, on the one hand, there is no oil dilution in the working chamber due to water condensate, and, on the other hand, there are no components at risk of corrosion. The cylinder bore can be designed with a ceramic base (SiC, etc.), and the carbon-based piston and piston rings are also not at risk of corrosion. The working chamber of the cylinder head contains no ferrous components such as the valves or valve seats of a four-stroke engine.

[0017] Despite the measures described, the two-stroke engine, even in its modern design, has considerable disadvantages, particularly with regard to exhaust emissions, which make its continued use rather questionable despite its undeniable advantages such as compact design, light weight, etc.

[0018] The present invention is based on the object of providing a method for operating a two-stroke internal combustion engine of the above-mentioned type, with which the harmful pollutant emissions can be kept particularly low in a particularly simple and reliable manner. Furthermore, a two-stroke internal combustion engine particularly suitable for implementing the method is to be provided.

[0019] With regard to the method, this object is achieved according to the invention in that, in a final phase of the exhaust stroke of the respective cylinder, the gas-side connection between the gas outlet window and the exhaust system is interrupted and instead the gas outlet window is supplied with fresh gas under excess pressure via the exhaust channel, preferably pure air as fresh gas.

[0020] Modern two-stroke engines often have multiple exhaust ports in the cylinder, which usually feed into a common exhaust system. Even if the text refers to only one exhaust port for the sake of simplicity, this explicitly refers to the total number of exhaust ports present in a cylinder and is intended to include them as well.

[0021] The invention is based on the consideration that a key reason for the adverse exhaust gas or emission values ​​of conventional two-stroke engines is scavenging losses and the associated comparatively low capture efficiency. With conventional port-controlled scavenging in a two-stroke engine, the exhaust opens first during the piston's expansion stroke to relieve the internal cylinder pressure by escaping exhaust gas (necessary pre-exhaust). Subsequently, the intake or transfer port(s) open, admitting fresh gas from the crankcase or, if applicable, from an upstream charger or compressor into the cylinder. Due to the design of the port-controlled two-stroke engine, however, the exhaust remains open longer than the transfer ports.Primarily in this time phase towards the end of the gas exchange, fresh gas in the cylinder enters the exhaust port via the still open gas outlet window, which causes the high proportion of fresh gas purge losses.

[0022] Typical valve timing for a slot-controlled two-stroke engine is typically between 120-190°CA (CA = crankshaft) in the exhaust area and between 90-140°CA in the fresh gas overflow area. This results in valve lengths for the pre-exhaust of approximately 15-35°CA. The long valve lengths are primarily intended for high-revving engines, whereas the short valve timing is designed for low operating speeds. To minimize the scavenging losses caused by these factors, the fresh gas entering the exhaust port should be backflushed into the cylinder. This improves the overall capture efficiency and allows the backflushed fresh gas to be fed into the regular combustion.

[0023] To achieve this, the section of the exhaust port adjacent to the cylinder's gas exhaust port is planned to be used as a temporary backflushing port. At the end of the gas exchange, the exhaust port is to be shut down prematurely and temporarily as an exhaust gas-carrying port and instead converted into an additional purge port for fresh gas. This is intended to push purge losses of air-fuel mixtures in the exhaust port back into the cylinder using pre-compressed fresh air.

[0024] Advantageous embodiments of the invention are the subject of the subclaims.

[0025] In an advantageous development, and particularly to increase engine efficiency, this backflushing should be accompanied by a boost or increase in charge density in the cylinder. To this end, a boost and / or increase in charge density of the gas volume in the cylinder is preferably achieved by supplying the pressurized fresh gas into the cylinder.

[0026] The final phase of the exhaust stroke, in which the gas-side connection of the gas outlet window to the exhaust system is switched to a gas-side connection of the gas outlet window to the fresh gas line carrying fresh gas under overpressure, is particularly preferably adapted to the design of the respective engine or cylinder with regard to the switching times. The start time for this final phase of the exhaust stroke is preferably selected at a time appropriately selected depending on the engine's operating point determined by load and speed, after the intake transfer ports open and before the intake transfer ports of the respective cylinder close.

[0027] This starting point can also depend on the combustion process (gasoline, diesel, HCCI, stratified charge, etc.) to be used, since the starting point for the backflushing of fresh gas also marks the end of the exhaust gas expulsion from the cylinder. The high temperature of the exhaust gas remaining in the cylinder is advantageous for controlling the ignition initiation for HCCI combustion processes (HCCI = homogeneous charge compression ignition). Therefore, a large spread of the starting point between 30° after the transfer port opens and the transfer port closes is advantageous for controlling such a combustion process.

[0028] For high engine loads or for a homogeneously operated gasoline engine, the starting point for the final phase of the exhaust stroke is preferably selected shortly before the transfer ports close. This is to fully utilize the scavenging cross-sections of the transfer ports and to prevent fresh gas already scavenged into the cylinder from being pushed back into the transfer ports. In addition, the end point for the aforementioned final phase of the exhaust stroke is advantageously selected to be the same as the end point of the exhaust process itself, so that the switching of the gas-side connection within the exhaust stroke is expediently completed when the gas exhaust window is closed by the piston moving in the cylinder.

[0029] The fresh air intended for backflushing and feeding into the exhaust port is most preferably provided at overpressure, ensuring reliable backflushing through the exhaust port into the cylinder, even under the prevailing pressure conditions against exhaust backpressure. To enable this, the fresh air can be pre-compressed or compressed in a suitably selected compressor unit, for example, an external supercharging 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 considered to be independently inventive, an exhaust-gas-driven charging device is provided for compression, in which, in particular, the enthalpy and / or the pressure energy carried in the exhaust gas stream is used to compress the fresh gas. Particularly preferred, and in a configuration 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, as is known, for example, from EP 2 846 019 A1, EP 2 846 020 A1, EP 3 061 970 A1, or EP 3 282 109 A1.

[0030] In this particularly preferred embodiment, which is considered to be independently inventive, the exhaust gas flowing out of the cylinder is fed, in a first phase of the exhaust stroke, entirely or partially to the primary side of an exhaust gas charging pump during the exhaust stroke, where it is further expanded to perform work. In a second phase of the exhaust stroke, the subsequently expanded exhaust gas located in the primary side of the exhaust gas charging pump is guided to the exhaust system together with any exhaust gas still present in the cylinder. Advantageously, the energy of the exhaust gas converted into expansion work in the exhaust gas charging pump on the primary side is converted, in whole or in part, on the secondary side into compression work of the fresh gas intended for feeding into the exhaust port.

[0031] Advantageously, by means of a suitable valve device in a first phase of the exhaust stroke at the beginning of the exhaust opening in the working cylinder, the exhaust gas flow in the exhaust duct is directed exclusively to the primary side of the charging device or the exhaust gas charging pump, in a second phase both the duct connection from the exhaust duct to the exhaust or exhaust system and the duct connection from the exhaust gas charging pump to the exhaust system are opened and in a third phase during the closing of the exhaust window in the working cylinder both the exhaust tract to the exhaust or exhaust system and the exhaust tract to the charging device are closed and at the same time the duct for fresh gas injection into the exhaust duct is opened.In an alternative advantageous embodiment, when using a turbocharger as the charging device, the second phase can also be omitted, since with this design, the exhaust gas from the cylinder can be completely routed through the turbocharger via the valve device. Diverting exhaust gas bypassing the turbocharger would at most serve the purpose of limiting the boost pressure via a so-called wastegate valve.

[0032] With regard to the two-stroke internal combustion engine, the aforementioned object is achieved by a multi-way switching valve installed in the exhaust port, via which the gas outlet port is connected both to the exhaust system and to a fresh gas overpressure line. This allows the backflushing of gas components that have entered the exhaust port, as provided for by the inventive concept, to be achieved in a particularly simple and reliable manner. Such a multi-way switching valve, on the one hand, can achieve the shutoff of the gas flow to the exhaust system, which is considered necessary for backflushing, so that from this point on, fresh gas present in the exhaust port between the cylinder and the valve device can no longer flow to the exhaust system as fresh gas loss.On the other hand, the fresh gas that is kept ready and under overpressure can be applied at the same time by the multi-way switching valve releasing the corresponding gas inlet into the outlet channel.

[0033] Advantageously, the multi-way switch valve can be used to establish a gas-side connection between the gas outlet port and the exhaust system, or between the gas outlet port and the fresh gas overpressure line. This allows the switch valve to easily temporarily close the exhaust duct to the exhaust system while simultaneously feeding the pressurized fresh gas into the exhaust duct.

[0034] For particularly reliable backflushing, depending on the individual design of the engine and its selected operating point, the volume of the exhaust port between the gas outlet window of the respective cylinder and the multi-way switching valve is advantageously at least 10% and / or at most 180% of the geometric displacement of the respective cylinder.

[0035] Exhaust gas aftertreatment (catalytic converter, particulate filter, etc.) can preferably be provided downstream of the valve device in the flow area of ​​the exhaust system. In the case of a turbocharger application, the exhaust aftertreatment would be located downstream of the turbine.

[0036] To provide the fresh gas under overpressure as intended by the design, the fresh gas overpressure line is advantageously connected on the inlet side to a compressor unit. This can be configured as an external supercharging unit, or the engine's own piston pump can be used for this purpose via the oscillating piston underside. In principle, all mechanically, electrically, or exhaust-powered compressors or conveying devices are suitable as a device for pre-compressing charge air, which can be used to backflush it into the cylinder through the exhaust port if necessary. In a particularly preferred embodiment, which is considered to be independently inventive, the compressor unit is formed by an exhaust-gas-driven supercharging device.

[0037] Preferably, a diaphragm pump according to the teachings of the aforementioned publications is provided for this purpose. This pump delivers and compresses fresh gas driven by exhaust pressure waves. The exhaust gas charge pump is characterized by very low exhaust backpressures, which are essential for the open charge cycle of the two-stroke engine, and, above all, by the utilization of a significant portion of the residual exhaust gas energy when opening the exhaust process.

[0038] In a further advantageous embodiment and for the purpose of a particularly efficient design, such an exhaust-gas-driven charging device is designed as a diaphragm pump, which is also controlled by the multi-way switching valve and is connected on the secondary side via the fresh gas overpressure line to the multi-way switching valve and on the primary side to the exhaust port. Most preferably, the multi-way switching valve is designed as a three-way valve, via which a gas-side connection can be established between the gas outlet port and the primary side of the compressor unit, if required.

[0039] Advantageous further process engineering embodiments in combination with the invention can be in particular: Fresh gas purging of the cylinder via the transfer ports can be achieved conventionally via the crankcase as a crankcase pump, optionally accompanied by mixed or separate lubrication of the crankshaft drive by adding oil to the fresh gas. The pre-compressed air from a compressor or turbocharger unit can either be used exclusively for backflushing via the exhaust port or can also fully or partially provide the fresh air supply for crankcase purging. Alternatively, the fresh gas supply to the transfer ports from the turbocharger unit can be achieved directly with compressed fresh air, without passing through the crankcase. This design of the cylinder purge air supply has the advantage of being able to completely decouple the crankshaft drive and its oil lubrication from the cylinder lubrication by installing an oil barrier between the crankshaft drive and the piston.The piston lubrication and crank chamber can be separated using a so-called crosshead crank drive, as is used, for example, in extremely long-stroke large engines. The crosshead itself, which among other things represents an oscillating linear bearing, is supplied with oil by the pressure lubrication of the crank drive and has an oil seal in the space below the piston, which on the one hand prevents lubricating oil from entering the lower piston chamber and on the other hand acts as a gas barrier between the crank chamber and the lower piston chamber. The space between the underside of the piston and the crosshead bearing varies its volume with the movement of the piston and can be used as an exclusive or additional fresh gas piston pump for the charge exchange. The fresh gas piston pump is solely responsible for pre-compressing the fresh gas, which is blown into the cylinder via the transfer ports in order to ensure, among other things, that the engine can be started.Alternatively, in exclusive use, the piston fresh gas pump can supply compressed fresh air in the piston's BDC region via the valve assembly to the exhaust port, flushing and recharging the cylinder accordingly. The additional use of the fresh gas piston pump with another compressor (compressor, turbo, exhaust gas charge pump, etc.) can be used, on the one hand, to increase the engine's air consumption; on the other hand, the fresh gas can be flushed into the cylinder via the transfer ports or the exhaust port of two separate compressors. The gas exchange of the fresh gas piston pump can be controlled by slots in the cylinder, diaphragm valves, throttle valves (partial load), or even by the valve assembly.In the head-valved four-stroke engine, the thermodynamically advantageous long-stroke design of the crankshaft drive reduces the scavenging cross-sectional areas (smaller valve diameters) in the cylinder head, which is detrimental to low-throttle gas exchange. In contrast, the port-valved two-stroke engine gains scavenging area with its long-stroke design thanks to larger port windows. Furthermore, the additional fresh gas scavenging via the exhaust port significantly increases the angular cross-sections on the scavenging side, resulting in advantageous fresh air cooling of the piston area on the exhaust side. As is generally known, the valve timing for gas exchange in a two-stroke engine is determined by the necessary time cross-sections of the scavenging windows and the required pre-exhaust to expand the residual gas in the cylinder.Increasing the time cross-sections for scavenging by means of the transformed exhaust port thus contributes to a shortening of the control times for scavenging and, accordingly, the overall exhaust process without reducing pre-exhaust. The effective strokes in both the expansion and compression strokes, which increase efficiency, are thus increased.

[0040] Further advantageous design embodiments in combination with the invention can be: To achieve high internal thermodynamic efficiencies with widely varying cylinder filling levels, variable geometric compression ratios are advantageous. This is particularly true for turbocharged gasoline engines, which must drastically reduce the compression ratio under high load to avoid knocking. Since there is no valve train in the cylinder head of a port-controlled two-stroke engine, a variable compression ratio can be implemented particularly easily by axially displacing (relative to the cylinder axis) the cylinder head or combustion chamber insert. The axial positioning of the cylinder head can be achieved mechanically, for example, by means of a threaded guide or eccentric shaft, but also hydraulically using a hydraulic piston.The sealing of the movable cylinder head / combustion chamber insert to the cylinder wall can preferably be carried out using self-lubricating (carbon) piston rings, since the adjustment travel is only a few millimeters and the sliding speeds are in the range of cm / s. The combustion chamber geometry of the modern two-stroke engine usually features a symmetrical constricted spherical cap with a central spark plug and circumferential crush ring gap. This combustion chamber design of the two-stroke engines enables considerably higher combustion speeds and thus less pre-ignition requirement and knock tendency compared to the roof-shaped combustion chambers in four-valve four-stroke engines. A particularly advantageous combustion chamber design with regard to high combustion speeds and knock resistance in conjunction with the exhaust port fresh gas scavenging process and optionally variable compression is achieved in combination with a so-called pre-chamber spark plug and a constricted spherical cap orOmega-shaped working chamber, which is primarily located in the cylinder head. The piston crown remains flat or only slightly curved. The high thermal load of an omega-shaped working chamber therefore does not occur in the piston area (similar to a diesel engine), but in the cylinder head, which has far better cooling capacity. For a variable compression ratio, it is advantageous to design only the central part of the cylinder-head-side combustion chamber (spherical or omega-spherical cap) to be movable and to leave the outer annular gap rigid, as this does not negatively increase the turbulence-generating squish gap height. The movable part of the cylinder-head-side combustion chamber would therefore be smaller in diameter than the cylinder bore. The lubrication process for the crankshaft drive and crosshead connecting rod guide via a closed circulation lubrication system on the one hand and the cylinder / piston lubrication on the other can be achieved using two different designs.On the one hand, conventional mixed lubrication using fuel-oil mixtures is possible for oil lubrication of the cylinder bore, although this requires a relatively high oil content. Classic separate lubrication, which uses a separate metering pump to deliver oil directly to the components to be lubricated (here, the cylinder bore) via oil lines, proves to be more advantageous in terms of relative oil consumption. A completely different tribological approach is offered by an encapsulated, oil-lubricated crankshaft, in which the cylinder bore is lubricated using dry lubricants, thus "oil-free." Experiments with carbon pistons and piston rings used in metal or ceramic cylinder bores are known from the literature. The lubricant used for this purpose is primarily based on graphite, i.e., pure carbon. Graphite has the advantage of remaining oxidation-stable up to 400°C, whereas oil begins to oxidize at around 200°C.During the expansion stroke of piston engines, the flame fronts also reach the cylinder bore, causing lubricating oil to evaporate from the cylinder bore and at least partially burn (CO emissions). This unavoidable phenomenon contributes both to increased pollutant emissions and to increased oil consumption in piston engines. Looking at future fuels, such as hydrogen, the effect of oil burnoff from the cylinder bore will be even more pronounced, as the hydrogen flame, due to its low ignition temperature, burns right up to the cylinder wall and the flame directly impinges on the lubricating film. In gasoline or alcohol-based petrol, the flame front extinguishes during the combustion phase at the latest a few hundredths of a millimeter from the cylinder wall, thus contributing to reduced oil burnoff.The tribological separation of crankcase oil lubrication and carbon-based solid lubrication for the cylinder area described here thus enables an "oil-free" combustion engine in the working cylinder area, resulting in the complete elimination of pollutants resulting from oil consumption and oil burn-off. In combination with hydrogen as a fuel, this would create an engine basis, purely from a design perspective, that produces no HC and hardly any CO emissions. Furthermore, the complete avoidance of contact between lubricating oil and the hot cylinder bore / combustion flame contributes to an increased service life of the lubricating oil, as its operating temperature is lower and no acidification occurs due to combustion residues introduced into the oil.The crosshead design of the crankshaft enables encapsulated forced oil cooling of the piston using pressurized oil from the crosshead bearing through the oil-carrying piston rod to the thermally highly stressed piston crown and piston ring area. The cooling oil also returns through the piston rod, so that no oil escapes below the piston crown, but flows exclusively back into the crankcase. The piston itself can be constructed from multiple parts, including different materials, such as carbon for the piston skirt and piston rings, or aluminum or steel for the piston crown / piston crown.

[0041] The advantages achieved by the invention are, in particular, that the intended process-related use of the exhaust port, in which it is shut down prematurely as an exhaust-carrying port at the end of the gas exchange and instead converted into an additional purge port for fresh gas, allows purge losses due to residual fresh gas in the exhaust port to be pushed back into the cylinder by means of pre-compressed fresh air and utilized for the upcoming combustion. This enables a drastic reduction in exhaust emissions, particularly those resulting from unburned fuel.Even in carburetor-driven, crankcase-scavenged two-stroke engines, this backflushing process can create a combustion-technically advantageous stratified charge in the cylinder at the end of the gas cycle, with high fuel concentrations near the spark plugs and relatively low-fuel zones in the piston crown area due to backflushing with fresh air. Essentially, the gas cycle is fundamentally transformed into a modified scavenging process with non-symmetrical timing via the symmetrical valve timing of the slot-controlled two-stroke engine.

[0042] The combination of the two-stroke engine with the exhaust gas charge pump, which is supplied with exhaust gas pulses from the cylinder on the inlet side and provides compressed fresh gas on the outlet side by utilizing the exhaust energy, also enables a particularly high level of engine efficiency. Overall, these design innovations can significantly positively influence the charge cycle, combustion, and expansion phase of the two-stroke engine, creating a highly efficient, low-emission, yet cost-effective piston engine for future applications, particularly with regard to range-extender drives and hybrid vehicles.This makes the well-known advantages of the two-stroke engine over four-stroke engines, such as small installation space requirements, high power density, low friction losses, good torque uniformity with a small number of cylinders and the like, available again for a wide range of applications, even in view of comparatively strict exhaust and emission regulations.

[0043] An embodiment of the invention is explained in more detail with reference to a drawing. In the drawing: Fig. 1 schematically shows a two-stroke internal combustion engine according to the prior art, Fig. 2 the control diagram of the engine according to Fig. 1 , Fig. 3 schematically shows an alternative two-stroke internal combustion engine according to the invention, Fig. 4 schematically shows the two-stroke internal combustion engine according to Fig. 3 during the exhaust stroke in an early, first phase, Fig. 5 schematically shows the two-stroke internal combustion engine according to Fig. 3 during the exhaust stroke in a second phase, Fig. 6 schematically shows the two-stroke internal combustion engine according to Fig. 3 during the exhaust stroke in a third or final phase, and Fig. 7 the control diagram of the engine after Fig. 3 .

[0044] Identical parts are provided with the same reference numerals in all figures.

[0045] The Fig. 1 The schematically illustrated combustion engine 1 is designed in a conventional manner according to the state of the art as a two-stroke engine. It comprises a number of cylinders 2, of which Fig. 1 only one is shown, 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 also act on a common crankshaft 8.

[0046] Within cylinder 2, in a conventional design, is the working chamber 10, in which a compressed fuel-air mixture is combusted during the cylinder's power stroke. In response, the working piston 4, which is movably arranged within cylinder 2, performs a power stroke, thereby driving the crankshaft 8. After completion of the power stroke, i.e., after expansion of the combusted working gas in cylinder 2 and reaching the so-called "exhaust opening," the combusted working gas is fed as exhaust gas during an exhaust stroke of cylinder 2 to an exhaust system 12 connected to the exhaust side of the cylinder.

[0047] For the gas exchange required to operate cylinder 2, the working chamber 10 is connected on the gas inlet side to the inlet transfer channels via a number of gas inlet windows 14 opening into the working chamber 10, and on the exhaust side to the exhaust system 12 via an exhaust channel 18 connected to the gas outlet window(s) 16 opening into the working chamber 10 (or, in the case of several gas outlet windows), or an exhaust channel system. The gas exchange in the working chamber 10 is controlled in the manner of a conventional two-stroke engine by the up and down movement of the working piston 4 in the cylinder, whereby the gas inlet windows 14 and the gas outlet window 16 are opened or closed as appropriate. In a conventional slot-controlled scavenging system in a two-stroke engine, the gas outlet window 16 and thus the exhaust opens first during the expansion stroke in order to relieve the internal cylinder pressure by escaping exhaust gas (necessary pre-exhaust).Subsequently, the gas inlet windows 14 open, and thus the transfer port(s) that admit fresh gas from the crankcase (or charger) into cylinder 2. The form of cylinder scavenging (reverse, cocurrent scavenging, etc.) is irrelevant, since in all processes the aim is to expel exhaust gas located in cylinder 2 by the inflowing fresh gas from cylinder 2 with the least possible mixing of both gas phases.

[0048] The sequence of the individual process steps during the gas exchange can usually be represented using a so-called control diagram 20, as is the case for the internal combustion engine 1 in Fig. 2 shown schematically by way of example. In such a control diagram 20, the angular position of the crankshaft 8 is plotted on the X-axis as a measure of the time during one complete crankshaft revolution in the unit °CA (= "degrees crankshaft) between the top dead center ("TDC") through the bottom dead center ("BDC") up to the top dead center again; the span between TDC and the next TDC is thus 360°CA. On the Y-axis, however, the degree of opening of the respective control unit is plotted, between "0" (for fully closed) and "1" (for fully open). In the control diagram 20 according to Fig. 2 The timing of the internal combustion engine 1 is shown for the degree of opening of the gas inlet window 14 and thus the gas inlet (line 22) as well as for the gas outlet window 16 and thus the gas outlet (line 24), which result from the degree of overlap by the working piston 4. In the example shown, the timing of the slot-controlled two-stroke engine 1 is between 120-190°CA (CA = crankshaft) in the exhaust range and between 90-140°CA in the (fresh gas) overflow range. This results in timing lengths for the pre-exhaust of approximately 15-35°CA. The long timing lengths primarily relate to high-revving engines, whereas the short timing is designed for low operating speeds.

[0049] Due to the design of such a slot-controlled two-stroke internal combustion engine 1, the exhaust typically remains open longer than the intake transfer ports. It is during this time phase toward the end of the gas exchange that the undesirable high proportions of fresh gas purge losses primarily occur, which subsequently enter the exhaust port 18 and from there into the exhaust system 12.

[0050] A modified two-stroke internal combustion engine 1' according to the invention, which overcomes this problem, is described in Fig. 3 shown. The two-stroke internal combustion engine 1' has a switchable multi-way switching valve 30 arranged in the exhaust channel 18, via which the gas outlet port 16 is connected both to the exhaust system 12 and to a fresh gas pressure line 32. Alternatively, a gas-side connection can be established via the multi-way switching valve 30 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.

[0051] Towards the end of the charge cycle, when the gas inlet windows 14 are already closed, the multi-way switching valve 30 blocks the gas flow from cylinder 2 towards the exhaust system 12. At the same time, the multi-way switching valve 30 closes the exhaust port 18 to the exhaust system 12, thereby opening the fresh gas overpressure line 32 to the exhaust port 18 on cylinder 2. This valve position is in Fig. 3 shown. In this valve position, pre-compressed fresh gas, preferably pure pre-compressed air, is forced into the exhaust port 18. This fresh gas pushes the gas mass present in the exhaust port 18, ie, the exhaust gas with the resulting fresh gas purge losses including fuel, back into cylinder 2 and simultaneously ensures recharging or charging in cylinder 2 at the end of the charge cycle.

[0052] With regard to the dimensioning of the components, the timing 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 power, while larger volumes reduce scavenging losses. The volume of the section of the exhaust port 18 located between the gas outlet window 16 of cylinder 2 and the multi-way switching valve 30 should preferably be at least 10% and / or at most 180% of the geometric displacement of cylinder 2.

[0053] 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 channel 18. In the exemplary embodiment, it is designed as a roller valve, in particular as a cylindrical, axially rotatable rotary valve in a stationary control housing. In the exemplary embodiment, it is provided with control cutouts that correspond to channel windows in the control housing. The rotary valve rotates synchronously with the crankshaft speed of the engine, whereby it can be driven by the engine indirectly via intermediate elements such as a toothed belt, chain, or gears, but also directly, in that the rotary valve is mounted directly on the crankshaft or forms part of the crankshaft. Alternatively, the drive of the (separate) rotary valve can also be electrical and thus mechanically separated from the crank drive.A phase adjustment between the rotary valve and the crankshaft position can also be provided to adapt the gas exchange, which can be influenced by the valve device, to the load and speed of the engine via variable phase position, particularly regarding the timing of the fresh air purge via the exhaust port. Individual control cutouts in the control housing can also be individually designed to vary their flow cross-section and phase position.

[0054] The rotary valve is preferably sealed to the control housing via a narrow gap, which offers the advantage of minimal friction. Cooling of the rotary valve is possible with liquid or air. The rotary valve itself can be designed as a hollow body with cooling medium flowing axially through its interior, and the rotary valve itself can be configured as a feed pump. A separate rotary valve is preferably mounted on rolling bearings.

[0055] In the Fig. 3 In the embodiment shown, the fresh gas overpressure line 32, in a configuration considered to be independently inventive, is in turn connected on the inlet side to the secondary side of an exhaust-gas-operated charging device 40, which thus forms a compressor unit for the fresh gas overpressure line 32. In this embodiment, the exhaust-gas-operated charging device 40, in a configuration also considered to be independently inventive, is designed as an exhaust gas charging pump, as is known, for example, from EP 2 846 019 A1, EP 2 846 020 A1, EP 3 061 970 A1, or EP 3 282 109 A1. For this purpose, it comprises a diaphragm pump 42, which is connected on the secondary side via the fresh gas overpressure line 32 to the multi-way switching valve 30 and on the primary side to the outlet channel 18.

[0056] For this purpose, the multi-way switching valve 30 is designed as a three-way valve, via which a gas-side connection can be established between the gas outlet window 16 and the primary side of the diaphragm pump 42 if required.

[0057] To clarify the operation of the components mentioned, the two-stroke internal combustion engine 1' is shown in FIG. Fig. 3 in the Figs. 4 bis 6 shown at different times during the exhaust stroke.

[0058] Fig. 4 shows the two-stroke internal combustion engine 1' during the exhaust stroke in an early, first phase, which serves for exhaust gas charging. In this first phase of exhaust gas charging, when the gas outlet port 16 opens, the channel connection to the exhaust system 12 is still completely closed by the valve position of the multi-way switching valve 30, and the exhaust gas is fed via the valve device exclusively to the primary side of the diaphragm pump 42 of the exhaust gas charging pump, where it expands through the conversion of the kinetic exhaust gas energy. This first phase of the charge cycle corresponds in function and timing approximately to the "pre-exhaust" of a conventional slot-controlled two-stroke engine. The control time, during which the exhaust port 18 is exclusively connected to the exhaust gas charging pump upon opening via the piston edge (exhaust gas charging), is preferably between approximately 10-50°CA.

[0059] In the subsequent second phase, according to the working position of the two-stroke internal combustion engine 1' acc. Fig. 5 , firstly, the gas inlet windows 14 open and thus the transfer ports in cylinder 2, and furthermore, the multi-way switching valve 30 opens the flow path from the exhaust port 18 to the exhaust system 12 and also the flow path of the relaxed, backward exhaust gas mass flow from the diaphragm pump 42 of the exhaust gas charging pump to the exhaust system 12. During this phase, the primary side of the exhaust gas charging pump is preferably permanently connected to the exhaust system 12 in order to enable the longest possible temporal relaxation of the exhaust gas in the exhaust gas charging pump. This corresponds to the operating mode fundamentally intended for the exhaust gas charging pump.

[0060] In Fig. 6 the two-stroke internal combustion engine 1' is in a third or final phase of the charge cycle, also according to the situation in Fig. 3 , shown. In this final phase, when the gas outlet window 14 in cylinder 2 begins to close due to the upwardly moving working piston 4, the multi-way switchover valve 30 closes both the flow paths between the exhaust port 18 and the exhaust system 12 and also those between the exhaust port 18 and the diaphragm pump 42 of the exhaust gas charging pump. In addition, a flow path is opened via the multi-way switchover valve 30, which now connects the exhaust port 18 to the fresh gas overpressure line 32 and thus to the secondary or fresh air side of the exhaust gas charging pump. As a result, compressed or pressurized fresh gas is pushed into the exhaust port 18, which in turn pushes the gas components in the exhaust port 18, primarily the scavenging losses from cylinder 2, back into the latter and also causes fresh gas recharging or supercharging in cylinder 2.

[0061] Finally, the gas outlet port 16 is closed by the ascending working piston 4, thus completing the gas exchange. Of course, several cylinders 2 can be supplied by an exhaust gas charging pump via one or more such valve devices.

[0062] The control diagram 50 of the two-stroke internal combustion engine 1' acc. Fig. 3 is in Fig. 7shown schematically. This shows the control times of the internal combustion engine 1' for the degree of opening of the gas inlet windows 14 and thus the gas inlet (line 52), for the gas outlet window 16 and thus the gas outlet (line 54), for the connection of the exhaust duct 18 to the exhaust system 12 (line 56), for the connection of the exhaust duct 18 to the fresh gas overpressure line 32 (line 58), for the connection of the primary side of the exhaust gas charging pump 42 to the outlet duct 18 during the exhaust pressure wave inlet into the exhaust gas charging pump 42 (line 60), and for the connection of the primary side of the exhaust gas charging pump 42 to the exhaust system 12 (line 62). List of reference symbols

[0063] 1.1' two-stroke internal combustion engine 2 cylinders 4 working pistons 6 connecting rods 8 crankshaft 10 working chamber 12 exhaust system 14 gas inlet port 16 gas outlet port 18 exhaust port 20 control diagram 22 line 24 line 30 multi-way changeover valve 32 fresh gas pressure line 40 exhaust gas-driven supercharging device 42 diaphragm pump 50 control diagram 52 line 54 line 56 line 58 line 60 line 62 line

Claims

1. Method for operating a two-stroke internal combustion engine (1') having a number of cylinders (2), in which a working piston (4) is respectively guided, and in which pressurized working gas can expand to perform work, wherein the or each cylinder (2) respectively has one or more gas exhaust windows (16), which are respectively connected to an exhaust system (12) via an exhaust channel (18), wherein in a final phase of the exhaust stroke of the respective cylinder (2), the gas-side connection between all gas exhaust windows (16) of the respective cylinder (2) and the exhaust system (12) is interrupted and instead all gas exhaust windows (16) of the respective cylinder (2) are supplied with fresh gas under pressure via the exhaust channel (18).

2. Method according to Claim 1, in which the supply of the fresh gas under pressure to the cylinder (2) produces charging and / or a charge density increase of the gas volume present in the cylinder (2).

3. Method according to Claim 1 or 2, in which the changeover of the gas-side connection provided in the final phase of the exhaust stroke is carried out at a start time within the exhaust stroke after the bottom dead centre of the piston position in the cylinder (2) and before the closing of gas inlet windows (14) of the respective cylinder (2), each of which is connected to an inlet overflow channel.

4. Method according to one of Claims 1 to 3, in which the changeover of the gas-side connection within the exhaust stroke is terminated when the gas exhaust windows (16) are closed by the piston (4) running in the cylinder (2).

5. Method according to one of Claims 1 to 4, in which, during the exhaust stroke, the exhaust gas flowing out of the cylinder (2) is supplied in whole or in part to the primary side of an exhaust gas charging pump (40) in a first stroke phase of the exhaust stroke and expanded there further to perform work before, in a second stroke phase of the exhaust stroke, the re-expanded exhaust gas present in the primary side of the exhaust gas charging pump (40) is guided together with any exhaust gas still present in the cylinder (2) to the exhaust system (12).

6. Method according to Claim 5, in which the energy of the exhaust gas converted into expansion work on the primary side in the exhaust gas charging pump (40) is converted in whole or in part on the secondary side into compression work of the fresh gas provided for injection into the exhaust channel (18).

7. Two-stroke internal combustion engine (1') having a number of cylinders, in which a working piston (4) is respectively guided, and in which pressurized working gas can expand to perform work, wherein the or each cylinder (2) respectively has one or more gas exhaust windows (16), which are respectively connected to an exhaust system (12) via an exhaust channel (18) characterized by a switchable multi-port switching valve (30) arranged in the exhaust channel (18), via which all gas exhaust windows (16) of the respective cylinder (2) are connected both to the exhaust system (12) and a fresh gas overpressure line (32).

8. Two-stroke internal combustion engine (1') according to Claim 7, in which a gas-side connection between the gas exhaust windows (16) and the exhaust system (12) or between the gas exhaust windows (16) and the fresh gas overpressure line (32) can alternatively be established via the multi-port switching valve (30).

9. Two-stroke internal combustion engine (1') according to Claim 7 or 8, in which the volume of the exhaust channel (18) between the gas exhaust window (16) of the respective cylinder (2) and the multi-port switching valve (30) is at least 10% and / or no more than 180% of the geometric swept volume of the respective cylinder (2).

10. Two-stroke internal combustion engine (1') according to one of Claims 7 to 9, the multi-port switching valve (30) of which is designed as a roller valve.

11. Two-stroke internal combustion engine (1') according to one of Claims 7 to 10, the fresh gas overpressure line (32) of which is connected to a compressor unit on the inlet side.

12. Two-stroke internal combustion engine (1') according to Claim 11, the compressor unit of which is designed as a mechanically or electrically powered compressor.

13. Two-stroke internal combustion engine (1') according to Claim 11, the compressor unit of which is formed by an exhaust gas-powered charging device (40).

14. Two-stroke internal combustion engine (1) according to Claim 13, the exhaust gas-powered charging device (40) of which is designed as a diaphragm pump (42), which is connected on the secondary side via the fresh gas overpressure line (32) to the multi-port switching valve (30) and on the primary side to the exhaust channel (18).

15. Two-stroke internal combustion engine (1') according to one of Claims 7 to 14, the multi-port switching valve (30) of which is a designed as a three-way valve, via which a gas-side connection can be established between the gas exhaust window (16) and the primary side of the compressor unit as needed.