EP4253738A1 - Method for operating a cyclically driven piston motor - Google Patents

Abstract

The present invention relates to a method for operating a clocked piston engine, which has at least one cylinder, at least one inlet valve arranged in the area of a cylinder head and at least one exhaust valve arranged in the area of the cylinder head and one in the at least one cylinder between a bottom dead center and a top dead center Has pistons that can be carried out by lifting movements. The process is intended to achieve an increase in efficiency. For this purpose, the method has the following steps: carrying out a charge change in the cylinder by means of the at least one inlet valve and the at least one exhaust valve, whereby a partial filling of the cylinder is achieved with a resulting compression or filling volume which is preferably predetermined by the valve control times, carrying out an extended expansion during the Downstrokes of the piston from top to bottom dead center, the expansion volume in the cylinder being preferably at least twice larger than the predetermined resulting compression or filling volume, releasing the exhaust gas after the expanded expansion from the cylinder by means of one exhaust valve or one of the exhaust valves Main exhaust gas cooling system, which releases cooled exhaust gas, preferably to the environment, by means of a controlled valve device and approximately isochoric cooling of the exhaust gas in the main exhaust gas cooling system at or up to a pressure that is below the ambient pressure, with cooled exhaust gas only then being released by means of the controlled valve device, when the pressure in the cylinder or in the main exhaust gas cooling system essentially reaches the ambient pressure or reaches a value that preferably deviates from the ambient pressure by a maximum of 0.5 bar. The invention further relates to a piston engine for carrying out this method.

Description

The present invention relates to a method for operating a clocked piston engine, which has at least one cylinder, at least one inlet valve arranged in the area of a cylinder head and at least one exhaust valve arranged in the area of the cylinder head, and one in the at least one cylinder between a bottom dead center and an upper one Dead center lifting movements executable piston.

Such a procedure is, for example, from DE 10 2008 014 249 A1 known. This document describes a method for changing combustion gases (purging) in a two-stroke internal combustion engine. Gas exchange occurs by means of a supercharging system with an inlet pressure level above the outlet pressure level, begins and ends in the upper half of the piston's stroke from bottom dead center (BDC) to top dead center (TDC). This only results in a partial filling of the cylinder because compression beyond the supercharged pressure level during the upward movement of the piston only takes place when the inlet valve is closed and the exhaust valve is closed at the same time. Since the inlet and outlet valves remain closed during the stroke movement from TDC to BDC, extended expansion occurs. In addition, there may be a control of the exhaust gas pressure level depending on the inlet pressure level. A throttle valve in the exhaust system can be used as an actuator to regulate the outlet pressure level.

• Durch eine Reduktion der Verlustwärme über die Brennraumoberflächen.
• Durch eine Reduktion der Verlustenergie im Abgas.

Some thermodynamic boundary conditions are explained below in order to better understand the beneficial effects of extended expansion. It follows from the first law of thermodynamics for open systems that, for a given amount of heat supplied (e.g. calorific value of the amount of fuel supplied), the technical work or efficiency of an internal combustion engine can only be significantly increased from a thermodynamic point of view by two measures:

• By reducing the heat loss via the combustion chamber surfaces.
• By reducing the energy loss in the exhaust gas.

• Es muss sich um ein ideales Gas handeln.
• Die Stoffwerte - insbesondere die Wärmekapazitäten - müssen als konstant angenommen werden.
• Es werden ideale Zustandsänderungen (adiabatische Kompression bzw. Expansion, ideale isochore, isobare oder isotherme Wärmezufuhr und Wärmeabfuhr) angenommen.
• Ein eventueller Ladungswechsel wird als Abkühlung ohne Massenaustausch simuliert.

The temperature-entropy diagram (Ts diagram) is usually used to estimate the efficiency of internal combustion engines. The Ts diagram needs some simplifying assumptions:

• It must be an ideal gas.
• The material values - especially the heat capacities - must be assumed to be constant.
• Ideal changes of state (adiabatic compression or expansion, ideal isochoric, isobaric or isothermal heat supply and heat removal) are assumed.
• A possible charge change is simulated as cooling without mass exchange.

Under these conditions, the Ts diagram shows the efficiency of different engine types, whereby the Ts diagram only takes into account the losses due to residual heat. Other losses (heat losses during compression and during or after combustion/heat supply), leakage and friction losses are not taken into account. The efficiency of a heat engine is defined as the quotient of the (technical) work dissipated to the thermal energy supplied. In the Ts diagram, the technical work can be read as the difference between the supplied and dissipated heat energy and thus as an area delimited by the cycle process.

The Fig. 1 shows schematically various cycle processes in the Ts diagram under the assumption that the heat is supplied up to the same maximum entropy (s).

The compression appears in the Ts diagram as an adiabatic change of state (1-2). When adding heat, a distinction is made between isochoric (2-3) and isobaric (2-3") entropy increase. There are also mixed forms with partly isochoric and subsequently isobaric entropy increase (2-2'-3'). This is also theoretical An isothermal entropy increase is conceivable (Carnot process), but this is difficult to implement in real machines and is therefore not shown.

The expansion is in turn an adiabatic change of state, depending on the degree of expansion, up to the isochore (3-4 or 3'-4 or 3"-4), up to the isobar (3-4' or 3'-4' or 3" -4') or up to the isotherm (3-4" or 3'-4" or 3"-4"). The larger the expansion volume - i.e. the lower the temperature after the adiabatic expansion - the lower the heat energy to be dissipated and the greater the efficiency.

Heat dissipation can take place along the isochore (4-1), along the isobar (4'-1) or along the isotherm (4"-1). A mixed form of heat dissipation is also conceivable. After expansion, for example up to the isobar (or even after an expansion to below the isobaric level), cooling can take place along the isochore (4'-5) to the isotherm followed by isothermal heat removal or compression (5-1) in order to increase the heat energy to be removed compared to the isobaric heat removal to reduce.

With the material properties of air and a compression ratio of 10:1, the following expansion volumes and efficiencies result for the above cycle processes (rounded): Description Line in the Ts diagram Compression volume VK Expansion volume VE Efficiency Equispace process (Otto cycle) 1-2-3-4-1 VK 1 × VK approx. 60% Seliger process 1 - 2 - 2' - 3' - 4 - 1 VK 1 × VK approx. 55% Diesel cycle *) 1 - 2 - 3" - 4 - 1 VK 1 × VK approx. 50 % Constant pressure process (Joule cycle) 1 - 2 - 3" - 4' - 1 VK 4xVK approx. 65% Cyclic process with isochoric heating, extended expansion and isobaric cooling 1 - 2 - 3 - 4' - 1 VK 4xVK approx. 70% Cyclic process with isochoric heating, extended expansion and isochoric and isothermal cooling 1 - 2 - 3 - 4' - 5 - 1 VK 4xVK approx. 75% Cyclic process with isochoric heating, extended expansion and isothermal cooling 1 - 2 - 3 - 4" - 1 VK 100 × VK approx. 85% *) Real diesel engines have a compression ratio higher than 10:1 and therefore also a higher efficiency.

From this table it can be seen that extended expansion to ambient pressure requires an expansion volume approximately 4 times larger than the compression volume. Extended expansion to ambient temperature requires an expansion volume about 100 times larger than the compression volume, which would result in very large displacements.

The required expansion volumes generally decrease as the compression ratio increases. For example, diesel engines or gas engines with high compression ratios require lower expansion ratios/expansion volumes than gasoline engines.

Furthermore, the required expansion volumes only apply to the theoretical case that no heat loss occurs during compression or during or after heat supply. In the real conditions of piston engines with internal combustion, in which significant heat loss occurs via the combustion chamber surfaces during and after combustion, the required expansion volume drops to approximately 3 times the compression volume (VE = 3 × VK).

Furthermore, the required expansion volumes for engines with internal combustion usually only apply to operation at full load.

The one from the DE 10 2008 014 249 A1 Although the known engine already works with extended expansion, the main focus of the control between the inlet pressure level and the outlet pressure level is to provide good flushing of the cylinder chamber, so that there is still a need for improvement in terms of efficiency.

The present invention is therefore based on the object of providing a method of the type mentioned at the outset with extended expansion, which has improved efficiency.

This task is solved by a method according to claim 1.

The procedure includes the following steps:
Carrying out a charge change in the cylinder by means of the at least one inlet valve and the at least one exhaust valve, wherein a partial filling of the cylinder is achieved with a predetermined resulting compression or filling volume, wherein preferably the resulting compression or filling volume is predetermined by the control times of the valves and therefore during can be changed during operation - for example by adjusting the phase of the camshaft,

Performing an extended expansion during the downward stroke of the piston from top to bottom dead center, wherein the expansion volume in the cylinder is, preferably at least 2 times, larger than the predetermined resulting compression or filling volume,

Discharging the exhaust gas after extended expansion from the cylinder by means of one exhaust valve or one of the exhaust valves into a main exhaust gas cooling system, which discharges cooled exhaust gas, preferably to the environment, by means of a controlled valve device, and

Cooling (preferably approximately isochoric) of the exhaust gas in the main exhaust gas cooling system advantageously at or up to a pressure which is advantageously more than 0.5 bar, but preferably at least 0.2 bar, below the ambient pressure, with cooled exhaust gas only then being cooled by means of the controlled Valve device is delivered when the pressure in the cylinder or in the main exhaust gas cooling system essentially reaches the ambient pressure or reaches a value which preferably deviates from the ambient pressure by a maximum of 0.5 bar and more preferably a maximum of 0.25 bar.

The aim of these measures is, on the one hand, that the heat dissipation takes place at a pressure that is advantageously below the ambient pressure, whereby this heat loss is minimized, and that, on the other hand, the push-out work is reduced, since the piston preferably pushes out the exhaust gas against a pressure lower than the ambient pressure must. The situation is optimal when the exhaust gas has cooled to ambient temperature when it is released from the main exhaust gas cooling system. This is why this is primarily done without further exhaust gas utilization (such as a turbocharger, etc.), because the necessary energy is already extracted from the exhaust gas. There is therefore an increase in efficiency due to the reduction in the heat to be dissipated after expansion has taken place. Due to the extended expansion and in particular the partial filling of the cylinder, an additional increase in efficiency can be achieved by reducing the time in which the hot gas can release heat via the cylinder or combustion chamber surfaces. In contrast to engines without partial cylinder filling, running through the constant space process requires significantly less than 360° crank angle, which means that the hot gas has significantly less time to release heat over the combustion chamber surfaces at the same speed. As a result, the constant space process (compression, heat supply and expansion up to the compression volume) has a higher efficiency in engines with a partial cylinder filling than in engines without a partial cylinder filling. This is followed by an increase in efficiency due to the extended expansion in engines with partial cylinder filling. In addition, the main exhaust gas cooling system has the advantage that cooling can take place over a longer period of time. As long as the associated exhaust valve is closed, part of the heat dissipation takes place independently of the processes and changes in state in the cylinder, whereby a crank angle of over 180° can be available for cooling and cooling down to ambient temperature can preferably be achieved more easily in the main exhaust gas cooling system can. An (electronically) controlled valve device can also Open when there is still negative pressure. This might even make more sense than opening at a pressure above ambient pressure. The "perfect" controlled valve device conveniently opens exactly when ambient pressure is reached and preferably opens a large cross section and then has minimal flow/throttle losses.

The method according to the invention can be used both for piston engines with internal heat supply and with external heat supply. With internal heat supply, the self-ignition and/or external ignition of a fuel or mixture takes place in the cylinder and with external heat supply, compression or pressure increase and heating takes place outside the cylinder.

In the present case, ambient pressure is understood to mean the pressure outside the main exhaust gas cooling system downstream of the controlled valve device, provided that the exhaust gas can no longer perform any further work that can be supplied to the engine or no significant cooling takes place. Ambient pressure is preferably understood to mean the pressure outside the “piston engine” system.

In an advantageous embodiment it is provided that the pressure in the cylinder is reduced below the ambient pressure due to the expanded expansion.

In a particularly advantageous variant, it is provided that the exhaust gas in the main exhaust gas cooling system is cooled essentially isochorically to a pressure below the ambient pressure in such a way that exhaust gas is ejected from the cylinder into the main exhaust gas cooling system primarily at a pressure that is lower than the ambient pressure. The term “isochoric” here means that during the cooling phase the exhaust gas is enclosed between the outlet valve and the controlled valve device, whereby a constant volume of the main exhaust gas cooling system can be assumed. The expansion volume can advantageously be designed so that the expanded expansion takes place at full load up to the ambient pressure, so the final pressure after the expansion at partial load is usually significantly lower than the ambient pressure. At partial load, expansion occurs into the negative pressure range, which significantly increases the efficiency at partial load. The primary goal of increasing efficiency is reduced heat dissipation after expansion, so that in any case heat dissipation occurs below the isobars that correspond to the ambient pressure (see 4'-1 in the Ts diagram of the Fig. 1 ), so that in the main exhaust gas cooling system and after opening the exhaust valve or valves there is also a lower pressure than ambient pressure in the cylinder during heat dissipation (negative pressure in relation to the ambient pressure). At a In an advantageous embodiment, the pressure should be more than 0.5 bar, but preferably at least 0.2 bar lower than the ambient pressure.

Alternatively, this negative pressure can also be suitable for supporting the gas exchange/flushing process instead of just reducing heat dissipation. This is particularly important in engines with internal combustion, which are preferably operated in the 2-stroke process, in which the gas exchange/scavenging takes place near bottom dead center or during the upward movement of the piston.

When opening one or more of the exhaust valves, the controlled valve device can advantageously separate the main exhaust gas cooling system from the environment at least as long as the pressure of the exhaust gas in the cylinder and/or in the main exhaust gas cooling system is substantially below the ambient pressure. It is therefore important to have optimal interaction between the exhaust valve and the controlled valve device in order to achieve a favorable increase in efficiency. The exhaust valve usually opens essentially at bottom dead center. The piston then moves upwards again. This causes exhaust gas to be actively pushed out of the cylinder, which also causes the pressure to rise again. As soon as ambient pressure is reached or the predetermined ratio is achieved, the controlled valve device opens and exhaust gas is expelled from the main exhaust gas cooling system without significant back pressure (because there is virtually no differential pressure). Due to the interaction according to the invention of the main exhaust gas cooling system and the controlled valve device, the (isochoric) cooling in the main exhaust gas cooling system takes place simultaneously but locally separately from the changes in state in the cylinder. This means that more time is available for heat dissipation at a given speed, and this enables (isochoric) heat dissipation and thus the creation of a negative pressure in the exhaust system (and thus a reduction in the heat loss to be dissipated to the environment).

In addition, an auxiliary exhaust system can be provided, into which at least a partial flow of the exhaust gas is introduced to reduce any excess pressure that may be present, which is essentially above the ambient pressure, and which preferably releases the partial flow of the exhaust gas, preferably into the environment, by means of a controlled valve device , whereby the pressure reduction essentially occurs before the remaining exhaust gas is discharged into the main exhaust gas cooling system. The auxiliary exhaust system therefore serves primarily to reduce the pressure in the cylinder if there is still excess pressure (above the ambient pressure) at the end of the extended expansion. This could occur particularly in the area of full load. In the auxiliary exhaust system, it is also preferred if the exhaust gas is finally released, in particular the environment only occurs after the partial flow of the exhaust gas has been subjected to cooling and thus a pressure reduction within the auxiliary exhaust system, so that a final expulsion of the exhaust gas into the environment also preferably only takes place essentially at ambient pressure. If there is negative pressure in the cylinder and in the auxiliary exhaust system, the controlled valve device should preferably also prevent an increase in pressure in the auxiliary exhaust system and in the cylinder due to the ambient pressure. The auxiliary exhaust system is preferably located behind a second exhaust valve and reduces any excess pressure in the cylinder near bottom dead center before a first exhaust valve in particular releases the connection to the main exhaust gas cooling system.

The advantages of the present invention have proven to be particularly effective when the piston engine is designed similar to a four-stroke piston engine with inlet and outlet valves arranged in the area of the cylinder head, but runs in two-stroke operation. This results in an increase in efficiency by reducing the energy for the gas exchange in engines that carry out the gas exchange at bottom dead center or during the upward movement of the piston, namely in two-stroke engines.

Furthermore, the one exhaust valve or one of the exhaust valves can remain open at least during the upward movement of the piston between bottom dead center and top dead center until the pressure in the cylinder or in the main exhaust gas cooling system reaches the opening pressure of the controlled valve device. This ensures that with each cycle there is a pressure increase in the cylinder and the main exhaust gas cooling system, so that exhaust gas is pushed out of the main exhaust gas cooling system, in particular into the environment. The exhaust valve conveniently opens near bottom dead center when there is negative pressure in the main exhaust cooling system due to isochoric cooling, but remains open long enough to allow exhaust gas to be expelled from the cylinder. In the variant with external heat supply, the entire exhaust gas is pushed out of the cylinder by the piston because the exhaust valve conveniently closes at or near top dead center. In the variant with internal combustion, the exhaust gas is partly pushed out by the fresh gas flowing in via the inlet valve or inlet valves and the closing of the outlet valve causes compression to begin.

Particularly in piston engines with internal combustion, the gas exchange advantageously takes place during the upward movement of the piston from bottom to top dead center with the inlet valve or one of the inlet valves open and the inlet valve or one of the outlet valves open, and then by closing the outlet valve or one of the outlet valves and the inlet valve or one of the inlet valves, the compression with the predetermined, resulting compression volume in the upper half of the piston stroke begins. A procedure is therefore used as already described in the DE 10 2008 014 249 A1 is known, with the present invention also providing a main exhaust gas cooling system. All aspects of the method and device described in this well-known document can also be used here, as far as makes sense; in particular, it is advantageous if the compression begins at the earliest at a crank angle of 60° before top dead center (ie, the inlet valve and outlet valve are closed) . The gas exchange then takes place approximately from bottom dead center to the closing of the exhaust valve. An additional charging of the gas introduced can preferably take place. The choice of fuel (gaseous or liquid) or mixture form is almost arbitrary. Depending on the type of application, external or self-ignition occurs. Preferably, the closing times and/or closing times of the one intake valve or the intake valves and/or the one exhaust valve or the exhaust valves can be changed in order to change the compression ratio.

However, the invention is also advantageous according to a process variant in which external heat is introduced. For this purpose, the gas, preferably air, can be compressed to a predetermined pressure by means of a compressor device outside the at least one cylinder before the gas exchange, then the compressed gas can be heated by means of a heating device and the compressed and heated gas can be fed in by means of the one inlet valve or one of the inlet valves Essentially during the downward movement of the piston from top to bottom dead center into the cylinder, a predetermined resulting filling volume (VK) which is smaller than this is achieved by closing the one inlet valve or one of the inlet valves, preferably in the upper half of the piston stroke Expansion volume (VE), and wherein preferably the closing times and / or closing times of the inlet valve or inlet valves can be changed in order to change the filling volume (VK). The expanded expansion then begins immediately after the intake valve is closed and occurs approximately adiabatic in the cylinder after the intake valve is closed. Heat can be introduced in any way (including through combustion). In particular, the design of a solar-powered air motor is conceivable.

In the variant with external heat supply, the filling volume or partial filling of the cylinder is determined by closing the inlet valve during the downward movement of the piston predetermined, and can be changed during operation via variable control times of the intake valve - for example by adjusting the phase of the camshaft.

Piston engines with external heat supply enable additional measures to increase efficiency. For this purpose, when the one inlet valve or one of the inlet valves is open, the gas can flow through the heating device and be heated in this as it flows through, preferably essentially at the pressure provided by the compressor device. This makes it possible for quasi isobaric heating to occur while the inlet valve is open because the gas flows through the heating device at a predetermined, preferably constant pressure. Due to the downward movement of the piston, a substantially isobaric expansion occurs as long as the intake valve is open, followed by an approximately adiabatic expansion in the cylinder after the intake valve is closed.

Advantageously, an increase in efficiency can be achieved in that, according to a method variant, the compressed gas is temporarily stored in an optional buffer storage and is supplied to the heating device with a pressure-controlled supply valve device and is heated essentially isochorically in the heating device as long as the inlet valve is closed. For this purpose, it is advantageous if the compressed gas volume is enclosed between the pressure-controlled supply valve device and the associated inlet valve and is exposed to the thermal effect of the heating device. More than 180° crank angles are available for this. It is therefore advantageous that compression and expansion are separated from each other. The (adiabatic) compression takes place in the compressor device. After the compressor device there is optionally an (isolated) buffer volume (provided by the buffer storage tank) which leads via an optional pressure-controlled supply valve device to an external combustion/heat supply heater device and then to the inlet valve of the cylinder. The timing of the intake valve (by closing the intake valve after, for example, a third of the piston stroke) causes the cylinder to be partially filled with subsequent expanded expansion (up to or below ambient pressure).

According to a further method variant, the processes can also take place in combination, in that the preferably approximately isochorically heated gas flows from the heating device into the cylinder with the inlet valve open or with one of the inlet valves open, then opens the controlled supply valve device when a predetermined pressure is reached, and more Gas from the buffer storage flows through the heating device and is essentially heated at the pressure provided by the buffer storage (i.e. approximately isobaric) and also flows into the cylinder. This can ensure, for example, that the filling pressure at least corresponds to the pressure from the buffer storage.

In addition, the invention also relates to a piston engine for carrying out the method according to one of claims 1 to 11. The piston engine is characterized by at least one cylinder, at least one inlet valve arranged in the area of the cylinder head, and at least one outlet valve arranged in the area of the cylinder head , a piston which can carry out lifting movements in the at least one cylinder between a bottom dead center and a top dead center and a main exhaust gas cooling system which is fluidly connected to the at least one exhaust valve and which can be separated from the environment by means of a controlled valve device in such a way that the exhaust gas is ejected from the cylinder into the main exhaust gas cooling system when the controlled valve device is closed at a first pressure level which is substantially below the ambient pressure, and that the exhaust gas is ejected from the main exhaust gas cooling system when the controlled valve device is open at a higher pressure level than the first pressure level. It is preferably a two-stroke piston engine, in which in the variant with internal combustion the purging/charge change preferably takes place during the upward movement of the piston with the inlet and exhaust valves (partially) simultaneously open and the cylinder is partially filled, and in the variant with external Heat supply preferably occurs when the cylinder is partially filled by closing the inlet valve(s) during the downward movement of the piston. The piston engine can be designed as a single-cylinder engine or as a multi-cylinder engine. In the version as a piston engine with internal combustion, the piston engine can be designed as a four-stroke engine (corresponding to a Miller engine, for example). However, in order to increase the power density, the design as a two-stroke engine is preferred. A considerable increase in efficiency can be achieved, particularly for large, slow-running piston engines.

According to a further advantageous embodiment, the one inlet valve or one of the inlet valves can have a variable flow cross section with increasing valve lift in such a way that flow into the cylinder initially occurs in the direction of the cylinder wall and, as the valve lift increases, increasingly in the direction of the center of the cylinder. Particularly due to the fact that the purging/charge change takes place during the upward movement of the piston with the inlet and exhaust valves (partially) open at the same time, this configuration is advantageous because this achieves a better complete charge change and prevents the fresh charge from overflowing into the exhaust system can be.

In addition, two inlet valves can be provided, one of the inlet valves being designed for the flow of fresh air into the cylinder and the other inlet valve being designed for the inflow of a, preferably gaseous, fuel or an ignitable mixture with a predetermined excess pressure.

When designing a piston engine with external heat supply, it is advantageously provided that a compressor device and a heating device are connected upstream of one inlet valve or one of the inlet valves, the compressor device and the heating device being designed in such a way that a substantially isobaric heat supply is provided at least temporarily by means of them can be effected in the cylinder. This is achieved in that gas delivered by the compressor device at a constant pressure is subjected to heat supply during the flow through the heating device. Of course, these are idealized perspectives; however, isobaric heat input takes precedence if losses due to other factors are not taken into account.

For this purpose, it can additionally be provided that an optional buffer storage and a controlled feed valve device are provided between the compressor device and the one inlet valve or one of the inlet valves, the controlled feed valve device at least temporarily separating the heating device from the compressor device and possibly the optional buffer storage and at When one inlet valve is closed or the inlet valves are closed, essentially isochoric heating is possible. This heating can occur during a significant period of the cylinder cycle (while the intake valve is closed), independent of other changes in state within the cylinder. An increase in efficiency can be achieved in this way.

When designing a piston engine with internal combustion, the second inlet valve can introduce the (gaseous) fuel or mixture into the cylinder, preferably with a certain excess pressure, with the second inlet valve closing shortly before the exhaust valve closes (start of compression without supercharging) or shortly after closing the Exhaust valve closes (start of compression with supercharging according to the excess pressure at the second inlet valve).

Furthermore, air can be blown tangentially into the inlet channel in front of the second inlet valve in order to create an air vortex in front of the inlet valve for good mixture formation, into which the (liquid or gaseous) fuel is injected.

Preferably, a (self-igniting) fuel, such as diesel, can be injected into the cylinder when the piston is near or shortly after top dead center.

Conveniently, a (gaseous) fuel can also be injected directly into the cylinder after purging with fresh air via the intake valve(s) and in which the corresponding mixture is spark-ignited when the piston is near or at top dead center. For this purpose, the injection process of the (gaseous) fuel can be completed shortly before the exhaust valve(s) close (which means that there is no charging before compression) or can take place shortly after the exhaust valve(s) close, which causes an increase in the pressure in the cylinder due to the pressure of the (gaseous) fuel takes place before compression (charging).

In a piston engine with external heat supply, the inlet valve can open when the piston is near top dead center and close when the piston is still in the upper half of the stroke, thereby allowing partial filling of the cylinder (and subsequently a extended expansion up to bottom dead center).

Fig. 1

ein Ts-Diagramm (mit unterschiedlichen Verfahrensabläufen),

Fig. 2

eine Prinzipskizze eines erfindungsgemäßen Kolbenmotors,

Fig. 3a-c

Skizzen zur Begriffsdefinition,

Fig. 4a-f

Skizzen zum Verfahrensablauf von einer ersten Kolbenmotorausführungsform,

Fig. 5a-b

Skizzen zum Verfahrensablauf von einer zweiten Kolbenmotorausführungsform,

Fig. 6a-b

Skizzen zum Verfahrensablauf von einer dritten Kolbenmotorausführungsform,

Fig. 7a-c

Skizzen zum Verfahrensablauf von einer vierten Kolbenmotorausführungsform,

Fig. 8a-e

Skizzengruppen zur Funktionsweise einer Einlassventilausführungsform, wobei jede Skizzengruppe eine schematische Draufsicht (oben links), einen schematischen Teilschnitt (unten links), und eine vergrößerte Ansicht des Teilschnittes (rechts) enthält,

Fig. 9a-c

Skizzengruppen zur Funktionsweise einer Ansaugtraktausführungsform, wobei jede Skizzengruppe eine schematische Draufsicht (oben) und einen schematischen Teilschnitt (unten) enthält,

Fig. 10

eine Prinzipskizze einer fünften Kolbenmotorausführungsform,

Fig. 10a-f

Skizzen zum Verfahrensablauf der fünften Kolbenmotorausführungsform,

Fig. 11

eine Prinzipskizze einer sechsten Kolbenausführungsform,

Fig. 11a-h

Skizzen zum Verfahrensablauf von einer sechsten Kolbenmotorausführungsform.

Embodiments of the present invention are explained in more detail below with reference to drawings. Show it:

Fig. 1

a Ts diagram (with different procedures),

Fig. 2

a schematic sketch of a piston engine according to the invention,

Fig. 3a-c

Sketches for the definition of terms,

Fig. 4a-f

Sketches of the process sequence of a first piston engine embodiment,

Fig. 5a-b

Sketches of the process sequence of a second piston engine embodiment,

Fig. 6a-b

Sketches of the process sequence of a third piston engine embodiment,

Fig. 7a-c

Sketches of the process sequence of a fourth piston engine embodiment,

Fig. 8a-e

Sketch groups for the functioning of an intake valve embodiment, each sketch group containing a schematic plan view (top left), a schematic partial section (bottom left), and an enlarged view of the partial section (right),

Fig. 9a-c

Sketch groups for the functionality of an intake tract embodiment, each sketch group containing a schematic top view (top) and a schematic partial section (bottom),

Fig. 10

a schematic sketch of a fifth piston engine embodiment,

Fig. 10a-f

Sketches of the process sequence of the fifth piston engine embodiment,

Fig. 11

a schematic sketch of a sixth piston embodiment,

Fig. 11a-h

Sketches for the process flow of a sixth piston engine embodiment.

The principle, based on Fig. 2 The explained structure of a piston engine 6 is similar to that of a four-stroke piston engine, except for the differences that will be explained below, and includes as essential components a cylinder 7, a crankcase 8, a crankshaft 9, a connecting rod 10, a piston 11, an exhaust valve 12, an intake valve 13, a spark plug 14 or a direct injection nozzle 15, and optionally an intake manifold injection nozzle 16 (petrol, etc.).

The present invention can be applied to various piston engine systems, which is why, for reasons of explanation, a mixed form containing the various variants is shown here, of which as a rule only one is used, which largely depends on the use of an internal or an external one Heat supply, ie, in particular the fuel used, etc., depends. The piston engine 6 can be designed as a single-cylinder engine or as a multi-cylinder engine. In the design as a piston engine 6 with internal combustion, the engine can be designed as a four-stroke engine (corresponding to a Miller engine, for example). However, in order to increase the power density, the design as a two-stroke engine is preferred. Therefore, only embodiments that are operated as two-stroke piston engines 6 will be described below.

In the design as a two-stroke engine with internal combustion, the purging or charge change takes place during the upward movement of the piston 11 with the inlet and outlet valves 12, 13 (partially) simultaneously open. This basically requires that for purging at least one inlet valve 13 there is a higher pressure than in the cylinder 7. A compressor device 19 required for this is only shown below if a particularly high flushing pressure is required. The following sketches for the respective process sequence show the piston positions (without connecting rods and crankshaft/crankcase) and the valve positions (without camshaft).

The invention is largely based on a piston engine 6 in which the expansion volume is significantly larger than the compression volume (preferably at least 2 times, better at least 3 times). This results in an extended expansion of the hot gas up to ambient pressure, almost to ambient pressure or into the negative pressure range. The expansion volume VE is generated by the total stroke of the piston 11 from top dead center (TDC) to bottom dead center (UT). The significantly smaller compression or filling volume VK is achieved by partially filling the cylinder 7.

In the embodiment as a piston engine 6 with internal combustion, this partial filling of the cylinder 7 is realized by a late start of compression in the upper half of the piston stroke (eg at approximately 60° before TDC). In the design as a piston engine with external combustion/heat supply, the compression takes place outside the cylinder 7 and the partial filling of the cylinder 7 takes place by closing the inlet valve 13/the inlet valves during the downward movement of the piston 11 in the upper half of the piston stroke (e.g. at approx. 60° after TDC). Based on Fig. 3a-c It can be seen that the following applies to the compression ratio for the version with internal combustion: Compression ratio = (VK + VV)/VV (e.g. 10:1).

The following applies to the expansion ratio: expansion ratio (VE + VV)/VV (e.g. 30:1).

This in Fig. 1 The Ts diagram shown (temperature-entropy diagram) of the piston engine 6 describes the changes in state of a gas packet as a closed cycle without mass exchange with the environment.

• (adiabatische) Kompression,
• (isotherme, isochore, isobare) Wärmezufuhr,
• (adiabatische) Expansion,
• (isotherme, isochore oder isobare) Wärmeabfuhr.

These state changes are:

• (adiabatic) compression,
• (isothermal, isochoric, isobaric) heat supply,
• (adiabatic) expansion,
• (isothermal, isochoric or isobaric) heat dissipation.

An essential core of the invention is that these state changes sometimes take place simultaneously, but are spatially separated.

The piston engine 6 according to the invention comprises, following the at least one exhaust valve 12, a main exhaust gas cooling system 17, at the output of which a controlled valve device 18 is arranged. The controlled valve device 18 is preferably a check valve which, in the preferred simplest version, opens when the ambient pressure inside the main exhaust gas cooling system 17 is reached.

The heat dissipation in the main exhaust gas cooling system 17 takes place permanently. In particular, this main exhaust gas cooling system 17 enables isochoric cooling of the hot exhaust gas as long as the exhaust valve 12/the exhaust valves of the piston engine 6 are closed. This isochoric cooling creates a negative pressure in the main exhaust gas cooling system 17, which also becomes effective in the cylinder 7 when the exhaust valve 12 is opened. If there is ambient pressure in the cylinder 7 and thus also in the main exhaust gas cooling system 17 due to the upward movement of the piston 11, the valve device 18 opens. With a suitable design of this exhaust gas cooling, the remaining heat dissipation occurs approximately isothermally, which minimizes the push-out work and the heat dissipation to the environment and thus the efficiency maximized.

Due to the inventive arrangement of the main exhaust gas cooling system 17 and the controlled valve device 18, the (isochoric) cooling takes place simultaneously, but locally separated, from the changes in state in the cylinder 7. This means that more time is available for heat dissipation at a given speed, and this results in isochoric cooling Heat dissipation and thus the generation of a significant negative pressure in the main exhaust gas cooling system 17 (and thus a reduction in the heat to be given off to the environment).

The following is based on the Fig. 4a-f a first embodiment of the piston engine 6 and the associated method are explained in more detail. It is a piston engine 6 in the design as a two-stroke engine with internal combustion, with an inlet valve 13 and an exhaust valve 12 per cylinder 7 and with the main exhaust gas cooling system 17 and the associated controlled valve device 18, preferably a check valve. The following symbols are used in the associated figures:

In Fig. 4a cylinder 7 is at circuit starting point 1. The inlet valve 13 and the outlet valve 12 are closed and the system is at the start of compression. The main exhaust cooling system 17 is located in section 4'-5. This means that the pressure in the main exhaust gas cooling system 17 drops below ambient pressure due to the exhaust gas cooling. Due to the closed outlet valve 12 and the closed valve device 18, isochoric cooling takes place.

When the piston 11 reaches top dead center (TDC), state point 2 is present (see Fig. 1 and Fig. 4b ). Depending on the type of internal heat supply, the fuel is injected (e.g. diesel) or the mixture is ignited. Fig. 4b refers to the most diverse variants of this internal heat supply, which is why it is listed here as an alternative the change of state from 2 to 3 (isochoric heat supply) or 2 to 3" (isobaric heat supply) or 2 to 2' to 3' (mixed heat supply) is referred to. In the main exhaust gas cooling system 17, the pressure drops further in the area 4'- due to the exhaust gas cooling. 5.

Referring to Fig. 4c The expanded expansion now takes place up to isobaric (ie from 3, 3' or 3" to 4'). Here, the expansion volume VE is approximately three to four times larger than the compression volume VK. Meanwhile, the pressure in the main exhaust gas cooling system 17 continues to fall isochorically 4' to 5.

By opening the exhaust valve 12 (see Fig. 4d ) near the UT, the negative pressure becomes effective in the cylinder 7 due to the previous isochoric cooling in the main exhaust gas cooling system 17. There is now negative pressure in cylinder 7 and in the main exhaust gas cooling system 17. The isochoric heat dissipation (4'-5) and in particular the subsequent isothermal cooling (5-1) now affects the cylinder 7 and the main exhaust gas cooling system 17 approximately equally.

Out of Fig. 4e It is now clear that the negative pressure in the cylinder 7 first reduces the push-out work (isothermal heat dissipation or compression 5-1). Furthermore, after opening the inlet valve 13, the flushing/charge change is supported. The valve device 18 only opens when ambient pressure is reached in the cylinder 7 or in the main exhaust gas cooling system 17.

In the next step ( Fig. 4f ) compression begins when the exhaust valve 12 and the inlet valve 13 close (circuit starting point 1). If the inlet valve 13 closes after the outlet valve 12, charging takes place with flushing pressure (as in Fig. 4f shown). In the main exhaust gas cooling system 17, after the exhaust valve 12 and the valve device 18 are closed, isochoric cooling (4'-5) begins again.

This type of engine is generally suitable for diesel engines because diesel (or other self-igniting fuels) ignites itself at or after TDC - so no mixture needs to be formed before or during compression.

For all fuels that very quickly form a good air-fuel mixture, especially for gaseous fuels such as methane or hydrogen, the fuel can alternatively be injected directly into the cylinder 7 during compression and the mixture formed in the cylinder 7 at or near TDC be ignited by means of a spark plug 14. If fuel is injected shortly before the exhaust valve 12 closes, no charging occurs. If the fuel is injected after closing the exhaust valve 12, charging occurs through the injection pressure of the fuel or through the injected fuel volume. Such a variant of a piston engine 6 is described below with reference to Fig. 5a-b shown. Since only the essential differences to the previous exemplary embodiment ( Fig. 4a-f ) is to be discussed, reference is made to the explanation of the previous method variant and engine type with regard to the further process that is not explicitly addressed here.

In Fig. 5a the cylinder 7 is located at the circuit starting point 1. The main exhaust gas cooling system 17 subjects the exhaust gas enclosed therein to isochoric cooling (4'-5). The inlet valve 13 and the outlet valve 12 are now closed and compression begins. (Gaseous) fuel is injected into the cylinder 7 from the injection nozzle 15. The injection pressure/fuel volume increases the cylinder pressure before or at the start of compression (supercharging). In Fig. 5b the piston 11 is at or near TDC (state point 2 in Fig. 1 ). The mixture is ignited. Depending on the process, an isochoric (2-3), isobaric (2-3") or mixed (2 -2'-3') heat supply is preferably carried out.

The embodiments from the Fig. 4a-f and 5a-b can make do with just one outlet valve 12 if the expansion volume (VE) is suitable for allowing the hot gas to expand up to or below ambient pressure.

Below is a third embodiment of the piston engine 6 according to the invention based on Fig. 6a-b explained in more detail. Reference should only be made to the most important differences from the previous exemplary embodiments, which is why the process flow with regard to Fig. 4a-f and 5a-b is referred to.

If the expansion volume (VE) is not sufficient to allow the hot gas to expand to ambient pressure and after the expansion there is still a pressure in the cylinder 7 that is above the ambient pressure, this embodiment provides a second outlet valve 12 ', which of expansion opens and reduces the cylinder pressure to ambient pressure before the first exhaust valve 12 opens. The second outlet valve 12' should open into an exhaust system, which can also be decoupled from the environment by means of a second controlled valve device 18'. Preferably, the second controlled valve device 18' is also a check valve, which preferably opens essentially at ambient pressure. This can prevent ambient air from flowing back into the cylinder 7, especially if there is already negative pressure in the cylinder 7 (e.g. at partial load). Furthermore, can An auxiliary exhaust gas cooling system 17' can also be connected to the second exhaust valve 12', so that a negative pressure can be generated in the cylinder 7 by opening the second exhaust valve 12' before the first exhaust valve 12 opens. The use of the described two outlet valves 12 and 12' as well as the main exhaust gas cooling system 17 and the auxiliary exhaust gas cooling system 17', each with controllable valve devices 18 and 18' connected thereto, is in the Fig. 6 from shown. Accordingly, after the expanded expansion, the residual pressure in the cylinder 7 near the bottom dead center (UT) is reduced by briefly opening the second exhaust valve 12 'to reach the state point 4' (see Fig. 1 ) to reach. Any excess pressure (above ambient pressure) that may be present is reduced by means of the valve device 18'. The main exhaust gas cooling system 17 is still in isochoric cooling during this period (see Fig. 6a ). Out of Fig. 6b It follows that by opening the first exhaust valve 12 near bottom dead center (BDC), the previous isochoric cooling in the main exhaust gas cooling system 17 behind the first exhaust valve 12 now takes effect in the cylinder 7. As a result, there is then a negative pressure in cylinder 7.

The remaining processes are the same as in the previous embodiments of the piston engine 6.

The embodiments described above are particularly suitable for fuels that only require a short time to form a mixture with air (e.g. diesel and gaseous fuels). For fuels that require a longer time for mixture formation (e.g. gasoline), a second inlet valve 13 'can be provided compared to the previous embodiments, which allows the cylinder 7 to be flushed with air during the upward movement of the piston 11, and the first inlet valve 13 can now have an injection nozzle 16 for the fuel (intake pipe injection). After purging with air via the second inlet valve 13', the mixture is introduced into the cylinder 7 via the first inlet valve 13 shortly before or shortly after closing the outlet valve 12/the outlet valves 12, 12'. For this purpose, on the one hand, the first inlet valve 13 must be subjected to a higher flushing pressure. For this purpose, a compressor device 19 is arranged in front of the first inlet valve 13. The mixture is already formed in the intake tract in front of the first inlet valve 13, preferably immediately after closing the first inlet valve 13 or near top dead center, in order to allow the longest possible time for mixture formation.

The gas routing on the two inlet valves 13, 13' should now fundamentally differ. At the second inlet valve 13 ', which is only responsible for flushing with air, an injection direction should be implemented, which reverse flushing of the entire cylinder 7 to the exhaust valve 12, so that all of the exhaust gas is displaced from the cylinder 7. At the first inlet valve 13, which introduces the mixture into the cylinder, an injection direction should be implemented which introduces the mixture into the cylinder 7 as centrally as possible.

• Gemischbildung von dem ersten Einlassventil 13, bevorzugt nahe des OT beginnend
• Spülung mit Luft über das zweite Einlassventil 13' während der Aufwärtsbewegung des Kolbens 11 bei gleichzeitig geöffnetem, ersten Auslassventil 12 (das in den Fig. 7a-c auch vorhandene zweite Auslassventil 12' zur Absenkung eines allfällig vorhandenen Restdrucks im Zylinder nach der erweiterten Expansion spielt für diese Erläuterungen keine Rolle).
• Einbringen des Gemischs mit erhöhtem Spüldruck kurz vor Schließen des Auslassventils 12 (keine Aufladung) oder kurz nach dem Schließen des Auslassventils 12 (Aufladung mit Spüldruck - wie in Fig. 7a-c dargestellt).

The thermodynamic processes are in principle the same as in the first embodiment ( Fig. 4a-f ). Therefore, only those processes are presented below in which a fourth embodiment (with two inlet valves 13, 13 ') differs from the first embodiment (with one inlet valve 13). The procedure briefly includes the following steps:

• Mixture formation from the first inlet valve 13, preferably starting near TDC
• Purging with air via the second inlet valve 13 'during the upward movement of the piston 11 while the first outlet valve 12 (which is in the Fig. 7a-c The presence of a second exhaust valve 12' to reduce any remaining pressure in the cylinder after the extended expansion also plays no role in these explanations).
• Introducing the mixture with increased flushing pressure shortly before closing the exhaust valve 12 (no supercharging) or shortly after closing the exhaust valve 12 (supercharging with flushing pressure - as in Fig. 7a-c shown).

The fourth embodiment therefore relates to a piston engine 6 in two-stroke operation with two inlet valves 13, 13 ', especially for fuels that require a longer time to form a mixture with air. In Fig. 7a It is shown that in the intake tract in front of the first inlet valve 13 - preferably near TDC - fuel is injected via an injection nozzle 16, so that a good air-fuel mixture can form until the first inlet valve 13 is opened. The main exhaust gas cooling system 17 (and also the auxiliary exhaust system 17') are, as always, in isochoric cooling (4'-5 in.) when the exhaust valves 12 and 12' are closed Fig. 1 ).

During the upward movement of the piston 11, flushing with air takes place via the second inlet valve 13 '( Fig. 7b ).

After flushing with air (via the first inlet valve 13, the mixture is released shortly before the outlet valve 12 closes (no boost) or shortly after the outlet valve 12 closes (charge with flushing pressure - as in Fig. 7c shown) introduced into the cylinder 7. The cylinder 7 is now at the circuit starting point 1, while the main exhaust gas cooling system 17 and, if like in the Fig. 7a-c present, also the auxiliary exhaust system 17 'in the isochoric cooling (from state point 4' to 5 in Fig. 1 ) condition.

The remaining processes are not shown any further because they correspond to the processes of the first embodiment ( Fig. 4a-f ), to which additional reference is made.

In piston engines 6 in two-stroke operation, in which the purging/charge change occurs during the upward movement of the piston 11 with the inlet valve 13 and exhaust valve 12 open (partially) at the same time, it is advantageous if the inflow direction at the inlet valve 13 at the beginning of the charge change is towards the cylinder wall or is not directed towards exhaust valve 12. In particular, in order to achieve a complete charge change and to avoid an overflow of the fresh charge into the main exhaust gas cooling system 17. As the gas exchange progresses, the inflow direction at the inlet valve 13 can and should also be directed centrally towards the cylinder axis. This can be realized by a special geometry on the inlet valve 13, which enables a variable flow cross section/opening cross section and thus a variable inflow direction depending on the valve lift.

This is explained below using the Fig. 8a-e explained in more detail. In Fig. 8a It can be seen that the valve seat 20 has a shoulder 21 towards the center of the cylinder. The paragraph 21 is beveled towards the cylinder wall. Fig. 8b shows the situation with the valve closed. Out of Fig. 8c it can be seen that the inlet valve 13 opens slightly. The open cross section (dark in the upper left images) on the inlet valve 13 points towards the cylinder wall. In Fig. 8d the inlet valve 13 is approximately half open. The opened cross section forms approximately a semicircle up to the valve axis. In Fig. 8e the inlet valve 13 is completely open. The opened cross section now corresponds to the opening cross section of the valve seat 20. Air/fresh gas can now flow in over the entire circumference of the associated inlet valve 13.

In piston engines 6 in two-stroke operation, in which the purging/charge change occurs during the upward movement of the piston 11 with the inlet valve 13 and outlet valve 12 (partially) simultaneously open, there is only a short time available for mixture formation at a given speed. It is therefore advantageous to provide an inlet valve 13 for introducing the mixture, in front of which the mixture is already formed before this inlet valve 13 opens. This is achieved by introducing air into the intake tract in front of the inlet valve 13, preferably tangentially. This creates an air vortex that is maintained even after the intake valve 13 is closed, and in which the fuel can be injected as early as possible (after the intake valve 13 is closed or close to TDC). One such an arrangement is in the Fig. 9a-c shown. In Fig. 9a It can be seen that an injection nozzle 16 for liquid or gaseous fuels is directed in the intake tract, preferably on the valve plate of the inlet valve 13. The air is supplied into the intake tract in front of the inlet valve 13 tangentially or offset to the axis of the intake tract. When the inlet valve 13 opens and the mixture enters the cylinder 7, the air flowing in creates an air vortex or swirl in the intake tract due to the tangential arrangement (see Fig. 9b ). After the inlet valve 13 is closed, this air vortex remains upright for some time. The fuel is preferably injected into this air vortex immediately after closing the inlet valve 13 or near TDC. Due to the air vortex and the earliest possible injection, the best possible mixture is generated in front of the inlet valve 13 at a given speed (see Fig. 9c ).

However, the present invention also relates to embodiments of piston engines 6 in which both external compression and external heat supply take place. In Fig. 10 a schematic diagram of a fifth piston engine embodiment with external compression and heat supply is shown. In front of the inlet valve 13 there is first a compressor device 19 and then an external heating device 22. The remaining components are known from the previous exemplary embodiments.

In the compressor device 19, the air is compressed adiabatically (circuit starting point 1 to state point 2 in Fig. 1 ). After the compressor device 19 there is approximately constant pressure. When the inlet valve 13 opens, the air flows through the external heating device 22 and is heated isobaric (state point 2 to state point 3 "in Fig. 1 ). In the main exhaust gas cooling system 17, the hot exhaust gas is permanently isochorically cooled with the exhaust valve 12 closed (from state point 4 'to intermediate point 5 in Fig. 1 ) (please refer Fig. 10a ).

As long as the inlet valve 13 is open, there is isobaric heating in the external heating device 22 or an isobaric expansion in the cylinder 7 (from state point 2 to 3 "in Fig. 1 ) (please refer Fig. 10b ).

Out of Fig. 10c It can be seen that the isobaric heating/expansion is completed by closing the inlet valve 13. The cylinder 7 is now filled with the filling volume VK, which corresponds to a partial filling of the cylinder 7. In cylinder 7 there is now status point 3" (see Fig. 1 ). The further expansion in cylinder 7 then takes place adiabatically (see Fig. 10c ).

The adiabatic expansion in cylinder 7 takes place up to bottom dead center UT (from state point 3" to 4' in Fig. 1 ). Alternatively, the adiabatic expansion can also extend into the negative pressure range (state point between 4' and 4" in Fig. 1 ) (see Fig. 10d ).

In the main exhaust gas cooling system 17, the exhaust gas was permanently isochorically cooled throughout the entire time (from state point 4' to 5 in Fig. 1 ). As a result, there is negative pressure in the main exhaust gas cooling system 17. By opening the exhaust valve 12 at bottom dead center UT, there is now also negative pressure in cylinder 7 (see Fig. 10e ), and initially approximately the same process takes place in cylinder 7 as in the main exhaust gas cooling system 17 (from state point 4 'to 5 in Fig. 1 ).

Out of Fig. 10f It can be seen that during the upward movement of the piston 11, the exhaust gas is pushed out into the main exhaust gas cooling system 17. Isothermal cooling/compression takes place in the main exhaust gas cooling system 17 (from intermediate point 5 to circuit starting point 1 in Fig. 1 ). Only when ambient pressure is reached in the cylinder 7 and in the main exhaust gas cooling system 17 shortly before top dead center TDC does the controlled valve device 18 open and release the cooled exhaust gas into the environment. With perfect cooling of the exhaust gas to ambient temperature, the smallest possible volume of the gas package to be expelled results, which at the same time minimizes the expulsion work and minimizes the heat to be dissipated (subisobaric heat dissipation according to the line 4'-5-1 in Fig. 1 ). Ideally, the gas to be pushed out then has the same volume as the inlet gas volume before compression began.

It should be noted that after opening the exhaust valve 12, conditions prevail in cylinder 7 and in the main exhaust gas cooling system 17, which are not exactly shown in the Ts diagram ( Fig. 1 ) can be represented, as they are spatially separated and sometimes run one after the other. For example, the temperature in cylinder 7 always corresponds to state point 4', since cooling only takes place in the main exhaust gas cooling system 17. Nevertheless, there is the same negative pressure in cylinder 7 as in the main exhaust gas cooling system 17. What is crucial for reducing the heat dissipated (and thus for maximizing efficiency) is that there is a negative pressure in cylinder 7 after opening the exhaust valve 12, and that the exhaust gas after Expansion in the cylinder 7 to the state 4 'while still in the system (with the valve device 18 closed) is preferably cooled to ambient temperature before it is pushed out into the environment.

In piston engines 6 of this type, the inlet valve 13 must seal against the pressure of the compressor device 19 or the pressure in the external heating device 22. In the Fig. 10a-f This is symbolized by an inverted inlet valve 13.

This embodiment makes it possible to minimize the heat energy to be dissipated after expansion has taken place up to isobaric (state point 4' in Fig. 1 ) or below the isobar - in extreme cases up to the isotherm (state point 4" in Fig. 1 ) by exhaust gas cooling down to ambient temperature and a valve device 18 after the main exhaust gas cooling system 17, which only opens when there is essentially ambient pressure in the cylinder 7 and in the main exhaust gas cooling system 17.

The following is based on the Fig. 11 and the Fig. 11a-h a sixth embodiment of a piston engine 6 is explained in more detail. This embodiment also enables heat supply above the isobars 2-3", which leads to an increase in the technical work corresponding to the area 2-2'-2" in the Ts diagram ( Fig. 1 ) so that the efficiency compared to the isobaric heat supply increases further. For this purpose, it is provided that there is a buffer storage 23 and an adjoining controlled feed valve device 24 (preferably a check valve) between the compressor device 19 and the external heating device 22.

This embodiment is based on the Fig. 11 briefly explained, with only the differences from the previous exemplary embodiment being discussed.

The following describes the procedure using the Fig. 11a-h briefly explained. According to Fig. 11a the air in the compressor device 19 is compressed adiabatically (from the circuit starting point 1 to state point 2 in Fig. 1 ). There is an approximately constant pressure in the buffer memory 23 (state point 2 in Fig. 1 ). Before the inlet valve 13 opens, isochoric heating takes place in the external heating device 22 (state point 2 after intermediate point 2 '). The air is virtually enclosed in the space between the controlled supply valve device 24 and the inlet valve 13.

After Fig. 11b prevails when the inlet valve 13 is opened at top dead center TDC, the condition in the cylinder 7 is as in the external heating device 22 after isochoric heating (intermediate point 2 'in Fig. 1 ). After Fig. 11c an approximately isothermal heat supply or expansion is carried out by a downward movement of the piston 11 in the external heating device 22 and in the cylinder 7 (intermediate point 2 'to intermediate point 2") until the pressure of the isobars (from state point 2 to 3" in Fig. 1 ) is reached.

When the pressure of the buffer storage 23 is reached in the cylinder 7 and in the external heating device 22 (corresponding to the isobars from state point 2 to 3 "in Fig. 1 ), the controlled supply valve device 24 opens between the buffer storage 23 and the external heating device 22. This is followed by an isobaric heat supply or expansion (intermediate point 2" to 3" in Fig. 1 ) (please refer Fig. 11d ).

By closing the inlet valve 13, the isobaric heating/expansion is completed. The cylinder 7 is now filled with the filling volume VK, which corresponds to a partial filling of the cylinder 7 (see Fig. 11e ). After closing the inlet valve 13, the expanded expansion in the cylinder 7 takes place adiabatically from (state point 3" to 4' in Fig. 1 ). When the inlet valve 13 is closed, isochoric heating begins again in the external heating device 22 (state point 2 after 2' in Fig. 1 ).

And as always, when the exhaust valve is closed, the hot exhaust gas is permanently isochorically cooled in the main exhaust gas cooling system 17 (from state point 4' to 5 in Fig. 1 ) The isochoric heating in the external heating device 22 with the inlet valve 13 closed and the isochoric cooling in the main exhaust gas cooling system 17 with the exhaust valve 12 closed take place spatially and temporally separately from the changes in state in the cylinder 7. As a result, a gas packet undergoes state changes in the Ts diagram that enclose a larger area. This area corresponds to the technical work of the cycle, therefore, according to the invention, the technical work and thus the efficiency are maximized for a given heat supplied.

At bottom dead center UT the adiabatic expansion is complete. The status point 4' now prevails in cylinder 7 (see Fig. 1 and Fig. 11f ). In the main exhaust gas cooling system 17, the hot gas was permanently isochorically cooled (state point 4' after 5 in Fig. 1 ). As a result, there is negative pressure in the main exhaust gas cooling system 17. By opening the exhaust valve 12 at bottom dead center (UT), there is now negative pressure in cylinder 7 (see Fig. 11g ), and initially approximately the same process takes place in cylinder 7 as in the main exhaust gas cooling system 17 (from state point 4 'to 5 in Fig. 1 ).

Out of Fig. 11h It can be seen that during the upward movement of the piston 11, the hot exhaust gas is pushed out into the main exhaust gas cooling system 17 and is cooled/compressed isothermally in the main exhaust gas cooling system 17 (from intermediate point 5 to circuit starting point 1). Only when ambient pressure is reached in the cylinder 7/in the main exhaust gas cooling system 17 shortly before top dead center TDC does the controlled valve device 18 open and discharge the cooled gas Exhaust gas volume into the environment. The heating in the external heating device 22 can take place in a variety of ways. Solar heating is also possible.

Reference symbol list

1

Circulation starting point

2

Condition after adiabatic compression

2'

Intermediate point after isochoric heat input

2"

Intermediate point after isothermal heat input from 2' to the isobar (2-3')

3

State point after isochoric heat input

3'

State point after isobaric heat supply following 2'

3"

Condition point after isobaric heat supply below 2

4

State point after adiabatic expansion to isochores (1-4)

4'

State point after adiabatic extended expansion to the isobar (1-4')

4"

State point after adiabatic extended expansion to isotherm (1-4")

5

Intermediate point after isochoric cooling from 4' to the isotherm (1-4")

6

Piston engine

7

cylinder

8th

crankcase

9

crankshaft

10

connecting rod

11

Pistons

12

outlet valve

12'

outlet valve

13

Inlet valve

13'

Inlet valve

14

spark plug

15

Injector (direct injection)

16

Injector nozzle (intake manifold injection)

17

Main exhaust cooling system

17'

Auxiliary exhaust cooling system

18

controlled valve device

18'

controlled valve device

19

Compressor device

20

Valve seat

21

Paragraph

22

external heating device

23

Buffer memory

24

controlled feed valve device

OT

Top Dead Center

UT

bottom dead center

VK

Compression volume

VE

Expansion volume

VV

Compression volume (remaining volume in the cylinder at TDC)

Claims (16)

Method for operating a clocked piston engine (6), which has at least one cylinder (7), at least one inlet valve (13, 13') arranged in the area of the cylinder head and at least one outlet valve (12, 12') arranged in the area of the cylinder head and one in which at least one cylinder (7) has pistons (11) which can carry out lifting movements between a bottom dead center (UT) and a top dead center (OT), the method comprising the following steps: Carrying out a charge change in the cylinder (7) by means of the at least one inlet valve (13, 13') and the at least one outlet valve (12, 12'), with a partial filling of the cylinder (7) with a resulting compression or filling volume is achieved, Carrying out an extended expansion during the downward stroke of the piston (11) from top to bottom dead center (OT, UT), the expansion volume (VE) in the cylinder (7), preferably being at least twice larger than the predetermined, resulting compression or filling volume ( VK), After the expanded expansion has taken place, the exhaust gas is released from the cylinder (7) by means of one exhaust valve (12) or one of the exhaust valves (12) into a main exhaust gas cooling system (17), which cools the exhaust gas by means of a controlled valve device (18), preferably to the environment, gives, and Cooling the exhaust gas in the main exhaust gas cooling system (17), wherein cooled exhaust gas is only released by means of the controlled valve device (18) when the pressure in the cylinder (7) and/or in the main exhaust gas cooling system (17) essentially reaches the ambient pressure or reaches a value , which preferably deviates from the ambient pressure by a maximum of 0.5 bar and more preferably a maximum of 0.25 bar. Method according to claim 1, characterized in that the exhaust gas in the main exhaust gas cooling system (17) is cooled substantially isochorically to a pressure below the ambient pressure in such a way that exhaust gas is ejected from the cylinder (7) into the main exhaust gas cooling system (17) primarily at a pressure occurs which is lower than the ambient pressure. Method according to claim 1 or 2, characterized in that when the one outlet valve (12) or one of the outlet valves (12) is opened, the controlled valve device (18) separates the main exhaust gas cooling system (17) from the environment at least as long as how the pressure in the cylinder (7) and/or in the main exhaust gas cooling system (17) is substantially below the ambient pressure. Method according to one of claims 1 to 3, characterized in that an auxiliary exhaust system (17 ') is provided, into which at least a partial flow of the exhaust gas is introduced to reduce any excess pressure that may be present, which is essentially above the ambient pressure, and this is preferably done by means of a controlled valve device (18') releases the partial flow of the exhaust gas, preferably to the environment, the pressure reduction essentially taking place before the remaining exhaust gas is released into the main exhaust gas cooling system (17). Method according to one of claims 1 to 4, characterized in that the piston engine runs in two-stroke operation. Method according to one of claims 1 to 5, characterized in that the one outlet valve (12) or one of the outlet valves (12) is open at least as long as the piston (11) moves upward between bottom dead center (UT) and top dead center (OT). remains until the pressure in the cylinder (7) or in the main exhaust gas cooling system (17) reaches the opening pressure of the controlled valve device (18). Method according to one of claims 1 to 6, characterized in that the gas exchange occurs during the upward movement of the piston (11) from bottom to top dead center (UT, TDC) with the inlet valve (13) or one of the inlet valves (13, 13') open. ) and the outlet valve (12) is open or one of the outlet valves (12, 12') is open, followed by closing the outlet valve (12) or one of the outlet valves (12, 12') and the inlet valve (13) or one of the Inlet valves (13, 13 ') the compression begins with the predetermined, resulting compression volume (VK) in the upper half of the piston stroke, and preferably the closing times and / or closing times of the one inlet valve (13) or the inlet valves (13, 13') and/or one or more of the exhaust valves (12, 12') can be changed in order to change the compression ratio. Method according to one of claims 1 to 6, characterized in that gas, preferably air, is compressed to a predetermined pressure outside the at least one cylinder (7) by means of a compressor device (19) before the gas exchange, then the compressed gas is heated by means of a heating device (22) and by means of the one inlet valve (13) or one of the inlet valves (13) the compressed and heated gas essentially during the downward movement of the piston (11) from top to bottom dead center (OT, UT ) is introduced into the cylinder (7), whereby by closing one inlet valve (13) or one of the inlet valves (13), preferably in the upper half of the piston stroke, a predetermined resulting filling volume (VK) is achieved, which is smaller than the expansion volume (VE), and wherein preferably the closing times and/or closing times of the inlet valve (13) or the inlet valves (13) can be changed in order to change the filling volume (VK). Method according to claim 8, characterized in that when the one inlet valve (13) or one of the inlet valves (13) is open, the gas flows through the heating device (22) and in this when flowing through, preferably substantially when flowing from the compressor device (19) provided pressure is heated. Method according to claim 8 or 9, characterized in that the compressed gas is temporarily stored in an optional buffer store (23) and is supplied to the heating device (22) by means of a controlled, preferably pressure-controlled, supply valve device (24) and in the heating device (22). Essentially isochorically heated. Method according to claim 10, characterized in that first the heated gas flows from the heating device (22) into the cylinder (7) with the inlet valve (13) open or with one of the inlet valves (13) open, then when a predetermined pressure is reached controlled supply valve device (24) opens and further gas from the buffer store (23) flows through the heating device (22) and is essentially heated at the pressure provided by the buffer store (23) and also flows into the cylinder (7). Piston engine (6) for carrying out the method according to one of claims 1 to 11, characterized by at least one cylinder (7), at least one inlet valve (13, 13 ') arranged in the area of the cylinder head and at least one outlet valve (12) arranged in the area of the cylinder head , 12 '), a piston (11) which can carry out lifting movements in the at least one cylinder (7) between a bottom dead center (UT) and a top dead center (TDC), and a main exhaust gas cooling system (17) which is flow-connected to the at least one outlet valve (12), which can be separated from the environment by means of a controlled valve device (17) in such a way that the exhaust gas is ejected the cylinder (7) into the main exhaust gas cooling system (17) with the controlled valve device (18) closed at a first pressure level which is essentially below the ambient pressure, and the exhaust gas is ejected from the main exhaust gas cooling system (17) with the controlled valve device (18) open ) occurs at a higher pressure level than the first pressure level. Piston engine (6) according to claim 12, characterized in that the one inlet valve (13) or one of the inlet valves (13, 13 ') has a variable flow cross section with increasing valve lift such that an inflow into the cylinder (7) initially in the direction which occurs on the cylinder wall and increasingly towards the center of the cylinder as the valve lift increases. Piston engine (6) according to one of claims 12 or 13, characterized in that at least two inlet valves (13, 13 ') are provided, one of the inlet valves (13') being designed for the flow of fresh air into the cylinder (11) and the other inlet valve (13) is designed for the inflow of a, preferably gaseous, fuel or an ignitable mixture with a predetermined excess pressure. Piston engine (6) according to claim 12, characterized in that one inlet valve (13) or one of the inlet valves (13) is preceded by a compressor device (19) and a heating device (22), the compressor device (19) and the heating device ( 22) are designed in such a way that a substantially isobaric supply of heat can be achieved at least temporarily. Piston engine (6) according to claim 15, characterized in that an optional buffer store (23) and a controlled feed valve device (24) are provided between the compressor device (19) and the one inlet valve (13) or one of the inlet valves (13), wherein the controlled supply valve device (24) at least temporarily separates the heating device (22) from the compressor device (19) and possibly the optional buffer storage (23) and, when one inlet valve (13) is closed or the inlet valves (13) are closed, a substantially isochoric heating occurs in the heating device (22) is possible.

Images