WO2011103868A2 - Moteur à combustion interne et procédé pour faire fonctionner un moteur à combustion interne - Google Patents

Moteur à combustion interne et procédé pour faire fonctionner un moteur à combustion interne Download PDF

Info

Publication number
WO2011103868A2
WO2011103868A2 PCT/DE2011/000189 DE2011000189W WO2011103868A2 WO 2011103868 A2 WO2011103868 A2 WO 2011103868A2 DE 2011000189 W DE2011000189 W DE 2011000189W WO 2011103868 A2 WO2011103868 A2 WO 2011103868A2
Authority
WO
WIPO (PCT)
Prior art keywords
heat exchanger
group
combustion engine
internal combustion
expander
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/DE2011/000189
Other languages
German (de)
English (en)
Other versions
WO2011103868A3 (fr
Inventor
Dieter Voigt
Ulrich Rohs
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
GETAS Gesellschaft fuer Themodynamische Antriebssysteme mbH
Original Assignee
GETAS Gesellschaft fuer Themodynamische Antriebssysteme mbH
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Priority claimed from DE102010047112A external-priority patent/DE102010047112A1/de
Application filed by GETAS Gesellschaft fuer Themodynamische Antriebssysteme mbH filed Critical GETAS Gesellschaft fuer Themodynamische Antriebssysteme mbH
Priority to CN2011800112317A priority Critical patent/CN102844541A/zh
Priority to DE112011100670T priority patent/DE112011100670A5/de
Publication of WO2011103868A2 publication Critical patent/WO2011103868A2/fr
Publication of WO2011103868A3 publication Critical patent/WO2011103868A3/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/02—Engines with reciprocating-piston pumps; Engines with crankcase pumps
    • F02B33/06—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L3/00—Lift-valve, i.e. cut-off apparatus with closure members having at least a component of their opening and closing motion perpendicular to the closing faces; Parts or accessories thereof
    • F01L3/08—Valves guides; Sealing of valve stem, e.g. sealing by lubricant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/02—Engines with reciprocating-piston pumps; Engines with crankcase pumps
    • F02B33/06—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps
    • F02B33/22—Engines with reciprocating-piston pumps; Engines with crankcase pumps with reciprocating-piston pumps other than simple crankcase pumps with pumping cylinder situated at side of working cylinder, e.g. the cylinders being parallel
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/44—Passages conducting the charge from the pump to the engine inlet, e.g. reservoirs
    • F02B33/443—Heating of charging air, e.g. for facilitating the starting
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M31/00—Apparatus for thermally treating combustion-air, fuel, or fuel-air mixture
    • F02M31/02—Apparatus for thermally treating combustion-air, fuel, or fuel-air mixture for heating
    • F02M31/04—Apparatus for thermally treating combustion-air, fuel, or fuel-air mixture for heating combustion-air or fuel-air mixture
    • F02M31/042—Combustion air
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M31/00—Apparatus for thermally treating combustion-air, fuel, or fuel-air mixture
    • F02M31/02—Apparatus for thermally treating combustion-air, fuel, or fuel-air mixture for heating
    • F02M31/16—Other apparatus for heating fuel
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L2810/00—Arrangements solving specific problems in relation with valve gears
    • F01L2810/05—Related to pressure difference on both sides of a valve
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies

Definitions

  • the invention relates to an internal combustion engine and a method for operating an internal combustion engine with intermittent internal combustion, comprising a compressor group consisting of at least one compression cylinder, at least one expander group consisting of at least two expansion cylinders and a heat exchanger arranged between the compressor group and the expander group in which a cold flow chamber of the heat exchanger has a working gas flow emerging from the compressor group and entering into the expander group, and a hot flow chamber of the heat exchanger has an exhaust gas flow emerging from the expander group.
  • the invention also relates to a method for operating an internal combustion engine with intermittent combustion and with a heat exchanger, wherein heat is transferred to a compressed in a compressor group working gas by means of the heat exchanger.
  • the invention relates to an internal combustion engine having at least one cylinder and at least one inlet-side charge exchange valve and at least one outlet-side charge exchange valve.
  • the invention relates to an internal combustion engine and a method for operating an internal combustion engine with at least one cylinder and at least one fuel line.
  • Combustion engines comprising a compression cylinder and a separate expansion cylinder, thereby enabling a four-stroke process in two spatially separated cylinders
  • US Pat. No. 5,857,436 describes an internal combustion engine having a compression cylinder and an expansion cylinder and a heat exchanger arranged therebetween, wherein exhaust gas heat which is not supplied to the working process is coupled back into the cycle via the heat exchanger.
  • a disadvantage of this design is the fact that only a limited space is available for the heat exchanger. This results in a limited efficiency of the internal combustion engine due to a caused by the charge change transient flow in the heat exchanger and because of a not inconsiderable effort to integrate the heat exchanger in the cylinder head of the engine.
  • an internal combustion engine with internal, intermittent combustion with one of at least one compression cylinder, preferably from at least two compression cylinders, existing compressor group, with at least one of at least two expansion cylinders expander group and one between the compressor group and the expander group arranged heat exchanger proposed, wherein a cold flow chamber of the heat exchanger exiting from the compressor group and entering the expander group working gas flow, wherein a hot flow chamber of the heat exchanger has a expelled from the expander exhaust gas stream and wherein the internal combustion engine is characterized in that the internal combustion engine downstream of the cold flow chamber has a working gas distributor with at least two outlets and the outputs of the working gas distributor in each case with an input of the expansion cylinder verb are unden.
  • the efficiency, in particular the thermodynamic efficiency, of the internal combustion engine via the coupling of the exhaust gas heat into the cyclic process can be increased by the advantage in an effective manner, which can be improved by a better and more continuous resp .
  • Quasi continuous gas flow through the heat exchanger due to the working gas distributor results.
  • the advantage of the continuous or quasi-continuous gas flow through the heat exchanger also brings advantages in the design of the heat exchanger, since transient processes can be largely minimized, whereby a heat exchanger suitable for this purpose can be developed and used with significantly less effort.
  • Internal intermittent combustion in this case describes combustion of a cylinder charge occurring within a cylinder and repeated in each cycle of the cylinder.
  • compression cylinder This internal, intermittent combustion usually takes place in generic combustion engines in the expansion cylinder, the expansion cylinder which implements the process steps “working” and “pushing out” of a four-stroke process.
  • the compression cylinder takes over the tasks “suction” and “compression” of a four-stroke process. Since a compression cylinder and an expansion cylinder need not necessarily be two cylinders adjacent to one another in an engine block, compression cylinders may be structurally arranged in a compressor group and the expansion cylinders structurally disposed in an expander group separate from the compressor group.
  • the internal combustion engine but also a method for operating an internal combustion engine, have an internal mixture formation.
  • the "internal mixture formation” here refers to the direct metering of fuel into a cylinder of the internal combustion engine. ⁇ br/> ⁇ br/>
  • a mixture formation within a closed cylinder is meant, after which the charge exchange valves of the respective cylinder are already closed after an intake process or after a charge change and no air mass flow into the cylinder.
  • an internal structure can be used Mixture formation are also attributed to the respective cylinder fuel when the exhaust valve is open, if in an exhaust gas flow through exothermic reactions heating of the exhaust gas is desirable. This embodiment can also be advantageous for a cold start.
  • a heat exchanger with separate material streams ie a heat-emitting material stream and a heat-absorbing material stream, is used as the heat exchanger in the above-mentioned context.
  • the transfer of heat takes place, according to the principles of thermodynamics, always from a hot flow of material to a cold flow of material, which is the colder working gas flow leading volume of the heat exchanger is defined as a "cold flow chamber” and the hotter exhaust gas flow leading and separated from the cold flow chamber volume Heat exchanger is defined as "hot flow chamber”.
  • This supply line can be, for example, a supply line from the cold-flow chamber to the expander group or else a supply line from the expander group to the hot-flow chamber.
  • the stroke volume of the compressor group may also correspond to the stroke volume of the expander group. It can thereby be realized a simple construction of the internal combustion engine, especially when an internal combustion engine is designed, for example, with two identical cylinder banks and between these cylinder banks space saving the heat exchanger is arranged.
  • an internal combustion engine with a heat exchanger is advantageous, which is characterized in that the compressor group has a smaller displacement than the expander group.
  • extended expansion is meant an expansion of the working gas or the exhaust gas beyond the intake volume, instead of an expansion up to the volume at the beginning of the compression or until the intake volume is increased fuel energy.
  • the expander group may have a higher number of cylinders than the compressor group.
  • the expander group may also have cylinders with a larger bore and / or cylinders with a larger stroke than the compressor group. This measure in turn brings the advantage of a higher efficiency by the previously described extended expansion.
  • the compressor group and / or the expander group have at least one charge exchange valve which opens in the flow direction.
  • Internal combustion engines according to the state of the art with gasoline or diesel processes almost exclusively have charge exchange valves, in particular poppet valves, with plates protruding into the combustion chamber, thereby forcing an opening of the valve into the combustion chamber.
  • the present in the combustion chamber combustion pressure causes a sealing effect, since the plate of the valve is pressed into its valve seat.
  • a charge exchange valve opening in the direction of flow has an opening direction pointing outward from the combustion chamber.
  • the compressor group or the expander group has at least one charge exchange valve opening counter to the direction of flow.
  • This proposal can be implemented advantageously when a higher pressure prevails in the inlet channel of a cylinder than in the cylinder itself, which may in particular be the case, for example, at an inlet valve of the expansion cylinders, when the pressure of the compressed working gas is applied to the closed inlet valve of the expansion cylinder the expansion cylinder expels or expels exhaust gas.
  • an outwardly opening valve may also have an automatic sealing action counter to the direction of flow.
  • the compressor group or the expander group may have at least one charge exchange valve designed as a slide.
  • a charge exchange control is hereby implemented, which is essentially independent of pressure.
  • the sealing effect of a charge exchange valve designed as a slide is only subordinate dependent on the applied gas pressures, since the opening direction and the direction of pressurization on the charge exchange valve diverge.
  • a charge exchange valve designed as a slide additionally offers the advantage of simple control or of a simple control drive, since the slide can be designed, for example, like a piston, whereby it is likewise controlled by the crankshaft.
  • the charge exchange valve has a crankshaft control.
  • the slide may in particular be a flat slide or a roller slide.
  • a simple design of the slider is possible, resulting in a cost advantage or even an advantage in the design.
  • an internal combustion engine which is characterized in that the expander group has an intake valve group and an exhaust valve group, wherein the intake valve group has a smaller flow area than the exhaust valve group.
  • the working gas passes under high pressure in the expansion cylinder and leaves this expansion cylinder after the expansion to above the volume at the start of compression or beyond the intake volume.
  • a relaxation of the working gas in the expansion cylinder takes place up to approximate ambient pressure. This results in an increase in volume, and consequently the volume flow through the outlet valve group can also be greater than the volume flow flowing through the inlet valve group.
  • two charge exchange valves of the intake valve group may be connected in series. This measure offers the advantageous possibility of providing a pressure relief at these charge exchange valves.
  • two charge exchange valves of the intake valve group connected in series have a different opening direction. It can be implemented on each of these charge exchange valves, the advantage of a self-contained sealing effect.
  • the first valve may have an outwardly directed opening direction, whereby a self-sealing valve seat is formed at a high pressure in the intake passage.
  • the second of these valves connected in series can consequently have an opening direction directed into the combustion chamber, whereby the combustion pressure in the expansion cylinder allows a further self-sealing valve seat.
  • the different sealing effect of both valve seats results in sum in a reliable sealed at each operating point valve group.
  • a camshaft actuating at least one charge-changing valve has a speed which is synchronous with the crankshaft speed.
  • An internal combustion engine with a compression cylinder and an expansion cylinder proposed according to this invention realizes in combination of both cylinders a four-stroke process in one crankshaft revolution instead of two crankshaft revolutions, whereby a two-stroke process per se is implemented in each cylinder.
  • control of the intake valve group and control of the exhaust valve group have different camshafts.
  • this can be used to further design the different camshafts for different cam forces for the intake valve group and the exhaust valve group.
  • one of these camshafts for particularly high cam forces can be designed and the other camshaft for lower cam forces, whereby the friction of the respective camshaft is present as needed and the mechanical efficiency of the engine is optimized accordingly.
  • a valve control of the compressor group and a valve control of the expander group may have different camshafts.
  • the expander group has means for internal mixture formation. If the fuel necessary for the combustion is supplied within an expansion cylinder, auto-ignition of the fuel-air mixture, for example in the heat exchanger, can advantageously be reliably avoided.
  • the internal combustion engine has a combustion process with auto-ignition.
  • a combustion process with auto-ignition can also be carried out as a diesel combustion process, with high efficiencies typical of this combustion process, so that it is proposed that the internal combustion engine have a heterogeneous fuel-air mixture when the autoignition conditions are reached.
  • the autoignition and the autoignition conditions result from the fuel used in each case.
  • a diesel burning process is not necessarily carried out with diesel fuel.
  • esterified vegetable oils, light oils, heavy oils, kerosene, gas oils, light gasoline, heavy gasoline, petroleum, synthetically produced fuels or combustible gases can be carried out.
  • an internal combustion engine with intermittent internal combustion a compressor group consisting of at least one compression cylinder, at least one expander group consisting of at least two expansion cylinders, and a heat exchanger disposed between the compressor group and the expander group wherein a cold flow chamber of the heat exchanger has an emerging from the compressor group and entering the expander group working gas flow and a hot flow chamber of the heat exchanger exiting from the Expander micher exhaust stream and wherein the internal combustion engine is characterized in that the compressor group has at least one automatically opening and automatically closing charge exchange valve ,
  • Such an automatically opening and automatically closing charge exchange valve has the advantage that under variable operating conditions a good control option, in particular the required compression end pressure, is given, so that by means of this solution also the efficiency of the internal combustion engine can be maximized.
  • An automatic control of the charge exchange valves can take place both in a compression cylinder and in an expansion cylinder.
  • the charge exchange valve may be an intake valve or an exhaust valve. This, in turn, has advantages for the efficiency, and to achieve maximum efficiency, an automatically opening and automatically closing valve at one point in the internal combustion engine in conjunction with a positively controlled valve at another point in the internal combustion engine combines the advantages of both types of control.
  • the charge exchange valve is designed as a plate, plate or ball valve.
  • this implements a simple and cost-effective design for the charge exchange valve.
  • an embodiment of an internal combustion engine having at least one self-opening and automatically closing charge exchange valve, preferably in the compressor group, is also advantageous independently of the other features of the present invention.
  • an internal combustion engine with internal intermittent combustion with a compressor group consisting of at least one compression cylinder, with at least one expander group consisting of at least two expansion cylinders and with a heat exchanger arranged between the compressor group and the expander group, is proposed. wherein a cold flow chamber of the heat exchanger has a leaving the compressor group and entering the expander group working gas flow and a hot flow chamber of the heat exchanger exiting from the Expanderenstein exhaust stream and wherein the internal combustion engine is characterized in that the heat exchanger is designed as a countercurrent heat exchanger.
  • This design offers the advantage of achieving the highest possible heat transfer efficiency even with limited installation space.
  • Such a heat exchanger may further be formed plate heat exchanger, whereby a structurally simple and cost-effective solution can be implemented.
  • a plate heat exchanger on its own advantageous for a generic internal combustion engine.
  • an internal combustion engine having intermittent internal combustion is arranged with a compressor group consisting of at least one compression cylinder, at least one expander group consisting of at least two expansion cylinders and one between the compressor group and the expander group Heat exchangers proposed, wherein a cold flow chamber of the heat transfer has exiting from the compressor group and entering the expander group working gas flow and a hot flow chamber of the heat exchanger exiting from the expander exhaust gas stream and wherein the internal combustion engine is characterized in that the hot flow chamber on the input side a higher number Connection bottles has as input side.
  • This embodiment of a heat exchanger particularly advantageously minimizes the surface responsible for the lost heat between the expander stage and the heat exchanger, in particular between the expander stage and the hot-flow chamber.
  • This embodiment also makes it possible to substitute a commonly used exhaust manifold or exhaust manifold, which in addition to the thermodynamic advantage in the increased efficiency additionally an advantage is achieved by reduced manufacturing costs.
  • the hot flow chamber has at least two connection flanges on the input side and a connection flange on the output side.
  • a heat exchanger for a generic internal combustion engine is advantageous if the heat exchanger has a connection flange for each expansion cylinder of the expander group.
  • the terms “input side” and “output side” are based on the flow direction of the exhaust gas flowing through the heat exchanger, the term “input side” meaning that side or piping or flange which is located between the heat exchanger Accordingly, the term “output side” designates that side of the hot flow chamber from which the exhaust gas exits the hot flow chamber again and flows into the downstream exhaust gas line and consequently into the environment.
  • an internal combustion engine with intermittent internal combustion having a compressor group consisting of at least one compression cylinder, at least one expander group consisting of at least two expansion cylinders, and one between the compressor group and the expander group Heat exchanger proposed, wherein a cold flow chamber of the heat exchanger exiting from the compressor group and entering the expander group working gas flow and a hot flow chamber of the heat exchanger has a expelled from the expander exhaust stream and wherein the internal combustion engine is characterized in that the heat exchanger is structurally located outside of the cylinder head.
  • design possibilities for the heat exchanger, in particular for its isolation can be made possible in a simple manner.
  • this design also has the advantage of keeping the cylinder head for a combustion engine according to the invention in its complexity to a small extent and thus optimally maintain design effort, development costs and durability in an optimal.
  • the heat exchanger can be arranged structurally outside the internal combustion engine. This also offers the advantage of every conceivable design option in To claim and decouple the heat exchanger structurally, vibrationally, thermally or acoustically from the internal combustion engine.
  • the heat exchanger is thermally insulated on an enveloping surface.
  • heat exchangers are constructed in a closed, gas-tight and usually made of metal housing. This housing, referred to as the "enveloping surface", because of the cold flow chamber and the hot flow chamber, is also exposed to an outward heat flow, resulting in a discharge of unusable energy, which reduces energy drainage to a minimum.
  • the heat exchanger has a second enveloping surface spaced from the first enveloping surface. This has the advantage of making the isolation particularly simple and without further aids available.
  • the space formed between the first enveloping surface and the second enveloping surface may be evacuated to at least eliminate the heat convection.
  • the "second enveloping surface” thus describes a cladding attached to the outer surface of the heat exchanger or a second housing resembling the heat exchanger housing, which is preferably also gas-tight.
  • the heat exchanger between the first enveloping surface and the second enveloping surface may include at least one heat-insulating material as the heat-dissipating substance, thus any material which controls the heat transfer from the inside the heat exchanger to the exterior of the heat exchanger adversely affected and thus causes an advantage in the energy balance of the internal combustion engine.
  • solids such as rock wool or glass wool, and g
  • Particularly effective and therefore advantageous for the energy balance of the internal combustion engine is a heat-insulating material with a heat exchanger adapted to the temperature. It is therefore proposed that the heat exchanger has exhaust gas between the first enveloping surface and the second enveloping surface. Exhaust gas which has left the hot flow chamber of the heat exchanger, thus, before it is conducted into the environment, the insulating gap of the heat exchanger can be supplied, whereby a particularly effective heat insulation is accomplished. Since a heat flow emerging from the heat exchanger is proportional to the temperature difference between the heat exchanger and the surroundings, the exiting heat flow is minimized by isolation by means of exhaust gas and the efficiency of the internal combustion engine is advantageously maximized. Cumulatively, it is also proposed that the heat exchanger between the first enveloping surface and the second enveloping surface from the interior of the heat exchanger has exiting exhaust gas.
  • the energy loss from radiant heat can also be countered by a reflective surface.
  • an internal combustion engine is proposed, even according to the above features, in which the heat exchanger has a main flow direction parallel to the cylinder bank.
  • the use of such an internal combustion engine in a motor vehicle is subject to the known restrictions of a limited space.
  • heat exchanger with a parallel to the cylinder bank main flow direction consequently offers the advantage to take the limited available space optimally and as needed. It is understood that such an arrangement of a heat exchanger is also advantageous independently of the other features in an internal combustion engine with a heat exchanger.
  • the heat exchanger has a connection flange with a maximum of 2.1 times the flange Nut height corresponding distance from the second enveloping surface.
  • the limitation of the distance of the connecting flange to the outer enveloping surface of the heat exchanger also provides independent of the other features of the present invention in an internal combustion engine with a heat exchanger the ability to realize the shortest possible hot tubing at the inputs or outputs of the heat exchanger.
  • connection flange and the outer surface of the heat exchanger are limited by the height of a flange nut used for mounting. If it is not necessary to mount a flange nut with the entire available thread on a mounting bolt, a distance less than 2.1 times the flange nut height between the connection flange and an outer surface of the heat exchanger is also conceivable.
  • the flange nut can be materially connected to the connecting flange.
  • a cohesive connection with the flange nut with the connecting flange is an easy to perform captive, is reduced by a flange immovably attached to the flange nut assembly work.
  • connection flange has a maximum distance of 1.1 times the flange thickness from the second enveloping surface if, in the case of assembly, the flange nut is already fastened to the connection flange and the bolt required for installation is subsequently introduced , This has the advantage that the inputs or the outputs, in particular the hot inputs of the heat exchanger, can be made even shorter and thereby even less loss of heat energy is made possible by a heat flow.
  • the connection flange has a thread for receiving a mounting means.
  • connection flange can have a maximum distance of 0.1 times the flange thickness from the further enveloping surface. This ensures that a bolt located in the flange is completely encompassed by the thread present in the flange and sufficient clearance remains between the end of the bolt facing the second enveloping surface.
  • connection flange is also advantageous, independently of the other features of the present invention, in an internal combustion engine having a heat exchanger connected to the internal combustion engine via at least one connection flange.
  • an internal combustion engine with intermittent internal combustion with a compressor group consisting of at least one compression cylinder, with at least one expander group consisting of at least two expansion cylinders and with a heat exchanger arranged between the compressor group and the expander group, wherein a cold flow chamber of the heat exchanger has a working gas flow exiting from the compressor group and entering the expander group, and a hot flow chamber of the heat exchanger has an exhaust gas flow emerging from the expander group and wherein the internal combustion engine is characterized in that a burner is arranged between the heat exchanger and the expander group.
  • a burner between the heat exchanger and the expander group offers the advantage, in the case of a cold start of the internal combustion engine, of reliably starting and also maintaining the combustion method used in this internal combustion engine if the heat exchanger is not yet in this operating state can transfer sufficient heat to the working gas. It is understood that a burner arranged between the heat exchanger and the expander group is located downstream of the cold-flow chamber, upstream of the cold flow chamber or upstream of the hot flow chamber may be arranged.
  • heat can be supplied to the working gas by means of a burner. This also advantageously a reliable cold start can be guaranteed. Likewise, this can in turn positively influence the firing process.
  • a variable air mass flow is supplied to the burner.
  • the supply of a variable air mass flow makes it possible to make variable the heat output supplied to the working gas and to go through the cold start to the desired operating state by means of a suitable control of the burner.
  • the burner can also be operated with a stoichiometric combustion air ratio. For a minimum installation space and a minimum fuel consumption, a maximum efficiency and a maximum burner output are also required for the burner. Maintaining a fixed combustion air ratio, preferably a stoichiometric combustion air ratio, can advantageously create the conditions required for this purpose.
  • the exhaust aftertreatment can have a burner. If the heat input in the exhaust aftertreatment by the combustion of fuel or other fuel in a burner, rather than by an exothermic reaction in the exhaust aftertreatment to heat the exhaust aftertreatment assembly to a required minimum temperature, this energy used by the Heat exchanger can be advantageously recycled to the cycle.
  • an internal combustion engine with intermittent combustion comprising a compressor group consisting of at least one compression cylinder, is provided with at least one of at least two expansion cylinders. henden expander group and proposed with a arranged between the compressor group and the expander heat exchanger, wherein a cold flow chamber of the heat exchanger has a emerging from the compressor group and entering the expander group working gas flow and a hot flow chamber of the heat exchanger has a leaving the expander exhaust gas stream and wherein the internal combustion engine characterized characterized in that the hot flow chamber is designed as an exhaust aftertreatment system.
  • a hot exhaust gas leaving the expander group can be advantageously used for exhaust aftertreatment, provided that the heat in the exhaust gas has not yet been completely or not completely removed by a heat exchanger.
  • a high temperature level is necessary or helpful because of the associated high reaction rates and catalytic effects.
  • a heat released by an exhaust gas aftertreatment can then also advantageously be fed back through the heat exchanger to the cyclic process.
  • the aftertreatment be effected by the application of heat. If an exhaust aftertreatment system located in the exhaust gas system has not yet reached a required minimum temperature, it may optionally be brought to a required temperature by means of a suitable heat input before the exhaust gas enters the heat exchanger and after the exhaust gas has exited the expander group.
  • the exhaust aftertreatment system may be a catalytic exhaust aftertreatment system. Cumulatively or alternatively, the exhaust gas aftertreatment system can also be an exhaust gas aftertreatment system acting as a particle filter and / or a nitrogen oxide oxide. be exhaustive exhaust aftertreatment system. All the above-mentioned embodiments allow an embodiment of the heat exchanger and an exhaust aftertreatment system, in which the advantages can be implemented synergistically. The above features may be used alternatively or cumulatively. It is understood that such an exhaust gas aftertreatment or such additional temperature control, in spite of an existing heat exchanger, in particular by a burner, is also advantageous for an internal combustion engine with heat exchanger independently of the other features of the present invention.
  • an internal combustion engine with at least one cylinder and at least one inlet-side charge exchange valve and at least one outlet-side charge exchange valve is proposed, which is characterized in that at least one charge exchange valve has a pressure compensation.
  • a charge exchange valve with a pressure compensation has the advantage that very high pressures in an inlet channel can be controlled and an undesirable opening of the charge exchange valve, in particular if this is closed only by a spring force, is effectively prevented at a pressure drop in the cylinder.
  • At least one charge-exchange valve on the intake side of an expansion cylinder can have a pressure compensation.
  • the charge exchange valve may have a plate and a cylinder-inside plate surface and the pressure equalization via a mounted on the shaft of the poppet valve compensating piston.
  • a backpressure can optionally be generated on this back side in a structurally simple manner by means of which it is possible to control pressure gradients in the inlet or outlet channels of the internal combustion engine.
  • Particularly advantageous is a given to the back of the double-acting piston pressure, which is applied shortly before opening the valve. With this measure, a pressure-balanced valve can be opened even at very high internal cylinder pressures, as a force acting on the back of the double-acting piston pressure can compensate the, an opening of the valve opposite, in-cylinder pressure.
  • the diameter of the balance piston may be larger than the diameter of the plate.
  • the diameter of the balance piston may be equal to the diameter of the plate.
  • the diameter of the balance piston may be smaller than the diameter of the plate.
  • valve train it may also be provided to apply a first contribution to the necessary closing force by a suitably designed valve spring and to apply the second contribution to the closing force via the compensating piston, which is necessary until the charge exchange valve is closed without leakage.
  • a suitably designed valve spring applies a first contribution to the necessary closing force by a suitably designed valve spring and to apply the second contribution to the closing force via the compensating piston, which is necessary until the charge exchange valve is closed without leakage.
  • the above considerations regarding the ratios of the diameters of the balance piston relative to the diameter of the plate may also be substituted by the effective effective areas.
  • the effective effective area of the balance piston is greater than the effective effective area of the plate, the effective effective area of the balance piston is equal to the effective effective area of the plate, or the effective effective area of the balance piston is small. ner is considered the effective effective area of the plate.
  • the term "effective effective area” designates the proportion of surfaces located on the compensating piston or plate, by means of which an adjacent gas force generates a resultant force in or against the opening direction or in the longitudinal direction of the charge exchange valve.
  • an internal combustion engine with internal, intermittent combustion with a compressor group consisting of at least one compression cylinder, with at least one expander group consisting of at least two expansion cylinders and with a heat exchanger arranged between the compressor group and the expander group
  • a cold flow chamber of the heat exchanger has a working gas flow exiting from the compressor group and entering the expander group
  • a hot flow chamber of the heat exchanger has an exhaust gas flow emerging from the expander group
  • the internal combustion engine is characterized in that the compressor group has compression pistons having a damaged volume of less than 5%, preferably less than 3%.
  • a deratig small Schadvolumen in the compression piston allows the highest possible compression of the working gas to be compressed, which advantageously a high power density of the internal combustion engine is achieved.
  • the damaged volume is not necessarily defined by the geometry of the compression piston, but can also be defined by the more precise configuration of the cylinder, in particular the cylinder space roof.
  • Cylinder space roof here means the area of the cylinder head delimiting the combustion chamber "referred to, as usual in the art, the reduced by the stroke volume of a cylinder maximum volume of a cylinder.
  • an internal combustion engine with at least one cylinder and with at least one fuel line is proposed, which is characterized in that the fuel line extends at least partially within an intake manifold.
  • an internal combustion engine with at least one cylinder and with at least one fuel line is proposed, which is characterized in that the fuel line extends at least partially within an exhaust manifold.
  • These embodiments make it possible, on the one hand, to heat the fuel in the fuel line, before it is fed to mixture formation, to a temperature level at which the mixture formation energy is considerably reduced and accordingly a mixture formation, for example within a combustion chamber, much faster and with a much lower Share of areas with heterogeneous fuel-air mixture takes place.
  • a much faster combustion initiation a much faster burn through of the charge and a much lower formation of soot particles take place.
  • the fuel may be heated to a temperature greater than 700 ° C, preferably to a temperature greater than 900 ° C, more preferably to a temperature greater than 1100 ° C.
  • the heating of the fuel to such a high temperature level has the advantage that the fuel is already decomposed in the fuel line or immediately after introduction into the working gas by radical formation or by other processes at the molecular level, whereby a required starting energy for the initiation of the combustion is significantly reduced and the combustion takes place at very high speed.
  • the fuel in an intake manifold of the internal combustion engine can be heated up.
  • the heating of the fuel by means of a likewise very hot working gas has the advantage that the fuel already has the temperature during introduction into the working gas and thus temperature gradients between fuel and working gas do not occur or are reduced to a minimum.
  • the fuel in an exhaust manifold mer of the internal combustion engine is heated.
  • the heating of the fuel in an exhaust manifold positively affects the efficiency of the internal combustion engine, since heat needed for heating would leave unused the accumulation sleeve placed around the internal combustion engine.
  • a heating of the fuel thus causes a heat recovery, since the amount of heat transferred to the fuel is recuperatively coupled into the cycle.
  • these solutions allow for a suitable embodiment of the method and the internal combustion engine, a waiver of additional heaters for the fuel.
  • the fuel line in which the preheating of the fuel takes place can be in particular a distributor tube of a storage injection system.
  • this brings with it the advantage that a synergistic effect arises in that already used and necessary components of the fuel system additionally act as heat exchangers.
  • this advantageously provides a common pressure line with an identical pressure and an identical temperature for all existing injection nozzles.
  • the "manifold" of the storage injection system refers to the memory of the storage injection system, ie the "rail" of a common rail injection system.
  • a further embodiment may additionally consist in that the internal combustion engine has means for heating a working gas upstream of the intake manifold. This is - as already explained above - further heating of the working gas to the desired working gas temperature allows, if the temperature of the working gas, for example, after a cold start or at an operating point with very little load, not the required temperature level of about 1 100 ° C has reached.
  • the "intake manifold” here refers to the piping supplying working gas to a cylinder and the working gas distributors therein, where "upstream” is the side of the intake manifold from which gas flows into this intake manifold.
  • the means for heating the working gas may be a burner.
  • a burner offers the advantage immediately after a cold start hot working gas to provide. It may be that only a partial oxidation of the working gas is required, so that further combustion within a cylinder can still take place.
  • means for heating consisting of an electric heating coil or a heat exchanger, whereby the working gas for combustion necessary oxygen is not withdrawn.
  • the "sidestream” thus refers to a working gas stream taken upstream of the intake manifold and, after passing the burner, is reintroduced into the intake manifold or optionally upstream of the intake manifold into the main stream of the working gas
  • the main flow and the bypass flow are not above their mass flows defined, so that the main flow can also lead to a smaller mass flow of working gas than the secondary flow.
  • the means for heating the working gas is an auxiliary heat exchanger, wherein an auxiliary cold flow chamber of the auxiliary heat exchanger has a working gas flow emerging from the compressor group and entering the expander group.
  • a heat exchanger used cumulatively or alternatively to a burner, the auxiliary heat exchanger can thus be used as a further starting aid, making use of an external heat supply.
  • a heat emerging from the internal combustion engine is recuperatively fed back via this auxiliary heat exchanger.
  • the auxiliary cold flow chamber preferably carries the main flow or the secondary flow of the working gas, and an auxiliary hot flow chamber may lead to exhaust gas.
  • the Hilfsh adoptedstrom- a chamber separate from the engine airflow, so an air flow, which no Working gas includes, leads and the working gas by means of an additional burner heat is supplied.
  • At least one expansion cylinder of the expander group may include at least one inlet valve closing 10 ° to 12 ° before top dead center of an expansion piston reciprocating in the expansion cylinder.
  • a very early closing of the exhaust valve causes a high residual gas content in the cylinder, this residual gas is compressed in the course of the working cycle.
  • the presence of an already precompressed charge in an expansion cylinder prevents, if the cylinder internal pressure equal to the pressure in an intake manifold of the expansion cylinder, opening an intake valve sventils exactly if the pressure before the valve equal to the pressure downstream of the valve and therefore no supercritical flow to the Valve occurs.
  • Supercritical flows near the speed of sound in a very narrow gap, such as may occur when a valve is opened, can result in excessive heating of the valve seat or valve.
  • undesirable flow conditions can adversely affect the combustion process by turbulence or shock waves.
  • FIG. 1 shows a first internal combustion engine with four cylinders in a schematic plan view with a two-cylinder compressor group, consisting of a two-cylinder expander and a heat exchanger;
  • FIG. 2 shows a combustion engine with five cylinders in a schematic plan view with a two-cylinder compressor group, consisting of a three-cylinder expander and a heat exchanger;
  • FIG. 3 shows an internal combustion engine with six cylinders in a schematic plan view with a compressor group consisting of two cylinders, with an expander group consisting of four cylinders and with a heat exchanger;
  • Figure 4 shows a six-cylinder internal combustion engine in a schematic plan view with a three-cylinder compressor group, with an expander group consisting of three cylinders and with a heat exchanger;
  • FIG. 5 shows an internal combustion engine with eight cylinders arranged in V-shape in schematic control with a compressor group consisting of four cylinders, with an expander group consisting of four cylinders, which are distributed in each case on both banks, and with a heat exchanger;
  • Figure 6 shows an internal combustion engine with eight arranged in V-shape cylinders in a schematic plan view with a four-cylinder compressor group, with an existing four-cylinder expander group, each in a
  • Bank are arranged, and with a heat exchanger
  • FIG. 7 shows an internal combustion engine with eight V-shaped cylinders in a schematic plan view with a compressor group consisting of three cylinders, with an expander group consisting of five cylinders and with a heat exchanger;
  • FIG. 8 shows an internal combustion engine in an arrangement corresponding to the internal combustion engine according to FIG. 6, wherein a burner is provided between heat exchanger and expander group and the arrangement is provided in the case of a cold start or cool heat exchanger;
  • Figure 9 shows a cross section through a cylinder head of an internal combustion engine, in particular also an internal combustion engine according to the figures explained above, with poppet valves and with a pressure equalization;
  • Figure 10 is a schematic view of a heated storage injection system for directly injected fuels, especially for an internal combustion engine according to the figures explained above;
  • Figure 1 a heat exchanger for an internal combustion engine, especially one
  • FIG. 1 shows an embodiment of an internal combustion engine 10 with four cylinders according to the invention in a schematic plan view with a compressor group 1 1 consisting of two compression cylinders 12 and an expander group 13 consisting of two expansion cylinders 14.
  • the respective compression cylinders 12 of the compressor group 11 suck in the working gas 15 from an environment and conduct the compressed working fluid. Gas 16 then into the heat exchanger 20.
  • the structural separation in a compressor group 1 1 and in an expander 13 here allows a compression process with less compressor work than is possible in a four-stroke process in common gasoline or diesel engines, since the compression cylinder 12 through this structural separation have cold cylinder tube and cold piston and thus the working gas to be compressed 15 is not additionally heated during the compression by a heat flow.
  • the existing temperature sink in the compression cylinders 12 rather leads to a cooled compression.
  • the working gas 15 is densified isothermally, with a minimum compressor work must be applied.
  • the compressed working gas 16 is then introduced into the cold flow chamber 22 of the heat exchanger 20 and transferred there into the cold stream 21 of the heat exchanger 20.
  • the cold stream present in this cold-flow chamber 22 undergoes a thermal state change by heating due to the hot stream 23, which is likewise located in the heat exchanger 20.
  • the present as cold stream 21 working gas in turn leaves the heat exchanger 20 as a heated working gas 17, which is supplied to the expander 13 in the further course.
  • the heated working gas 17 is evenly distributed to the expansion cylinder 14 of the expander 13.
  • the combustion method in this embodiment mainly includes a diesel process, which can be preferably used by the heated working gas 17 supplied to the expansion cylinders 14.
  • the working gas 17 is replaced by the heat in the heat exchanger 20 is a particularly high thermal energy, so that in the heated working gas 17 conditions for auto-ignition of a fuel in the expansion cylinder 14 are present.
  • a further compression in the expansion cylinders 40 is conceivable. Consequently, the presence of auto-ignition conditions in the heated working gas 17 is not necessarily required for this combustion process.
  • the hot exhaust gas 18 emerging from the expander group 13 is finally passed again into the heat exchanger 20, in which it is transferred into the hot stream 23.
  • the hot stream 23 is passed through the hot flow chamber 24 and there outputs its not converted during the expansion heat energy to the cold stream 21 from.
  • the exhaust gas leaves the heat exchanger as a cold exhaust gas 19 and can then be discharged in turn into the environment.
  • FIG. 2 Another internal combustion engine 10 with 5 cylinders according to the invention is shown in FIG. 2 in a schematic plan view with a compressor group 11 consisting of two compression cylinders 12 and an expander group 13 consisting of three expansion cylinders 14.
  • the compression cylinder 12 and expansion cylinder 14 each include the same displacement, whereby the larger by one cylinder Expanderffle 13, the heated working gas 17 has been expanded under a larger pressure ratio than the working gas 15 has been compressed in the compressor group 1 1.
  • This measure increases the output in the Expanderussi 13 piston work and leads to a further increasing efficiency over internal combustion engines with common diesel or Ottobrennclar.
  • FIG. 1 and 2 of the internal combustion engine 10 An essential feature of the design shown in Figures 1 and 2 of the internal combustion engine 10 is the arrangement of the heat exchanger 20, whose longitudinal extent parallel to the crankshaft axis of the internal combustion engine 10 extends.
  • This advantageous arrangement of heat exchanger 20 and internal combustion engine 10 makes optimum use of the space available in a vehicle.
  • Schematically, the design of the heat exchanger 20 is shown, which is performed from cross-countercurrent heat exchanger.
  • any other design for a heat exchanger 20, preferably types with a characteristic longitudinal extension, such as shell and tube heat exchangers can be used.
  • FIG. 3 shows a six-cylinder internal combustion engine 10 in a schematic plan view.
  • the embodiment illustrated in FIG. 3 has a compressor group 11 consisting of two compression cylinders 12 and an expander group 13 consisting of four expansion cylinders 14.
  • the principle of extended expansion is applied to a greater extent than would be possible in the embodiment according to FIG.
  • FIG. 4 shows an embodiment with a compressor group 11 consisting of three compression cylinders 12 and an expander group 13 also consisting of three expansion cylinders 14.
  • the principle of the extended expansion can also be realized here by the compressor cylinders 12 having a smaller stroke or have a smaller bore than the expansion cylinder 14.
  • compression cylinders 12 and expansion cylinders 14 can also be arbitrarily selected in a crankcase of an internal combustion engine 10.
  • the arrangement of two compression cylinders 12 or more expansion cylinders 14 as adjacent cylinders makes it possible to design the cooling of the respective expander group 13 or compressor group 1 1 as required.
  • FIG. 5 shows an internal combustion engine 10 shown with eight cylinders, wherein the respective compression cylinder 12 and expansion cylinder 14 are evenly distributed on both cylinder banks of the engine 10 and thus two compressor groups 11 and two expander groups 13 are present.
  • the heat exchanger 20 can be arranged in a space-saving manner between the two cylinder banks.
  • the hot exhaust gas 18 in particularly short distances in the heat Transfers 20 led.
  • the heat exchanger 20 bundles the hot stream 23 and leads the cold exhaust gas 19 in a common exhaust line.
  • the working gas 15 is directed on two separate sides of the internal combustion engine 10 in the respective compressor group 11 of the various cylinder banks.
  • the heat exchanger 20 bundles the compressed working gas 16 and leads the working gas 17 heated in the heat exchanger 20, in turn, into a plurality of working gas distributors, whereby the expansion cylinders 14 of the expander groups 13 are uniformly charged with heated working gas 17.
  • FIG. 6 of an internal combustion engine 10 with a compressor group 11 and an expander group 13 includes the compression cylinders 12 and the expansion cylinders 14, each in a separate cylinder bank.
  • This design makes it possible to design the thermal management, ie the cooling strategy of the compressor group 11 or the expander group 13, to the respective thermal conditions.
  • This arrangement also offers the advantage that the extended expansion can be realized by a different bore between the compression cylinders 12 and expansion cylinders 14, since the compression cylinders 12 are housed in a separate cylinder bank.
  • the arrangement of all compression cylinders 12 in a common cylinder bank also causes the working gas 15 is guided in each case in a common pipe string.
  • FIG. 7 A further embodiment with extended expansion is shown by the internal combustion engine 10 shown in FIG. 7.
  • This view shows the internal combustion engine 10 with eight V-shaped cylinders in a schematic plan view with a compressor group 11 consisting of three compression cylinders 12 and one out of five Expansion cylinder 14 existing expander group 13, wherein an expansion cylinder 14 is arranged in a cylinder bank together with the compression cylinders 12 of the compressor group 1 1.
  • the stream of the working gas 15 introduced into the compressor group is also subject to compression in the compression cylinders 12 with a pressure ratio which is less than the pressure ratio during the expansion of the heated working gas 17 in the expander group 13, which in turn enables thermodynamically better efficiency becomes.
  • the internal combustion engine 10 shown in FIG. 8 corresponds to an arrangement corresponding to the internal combustion engine 10 according to FIG. 6, a burner 40 being provided between the heat exchanger 20 and the expander group 13, and the arrangement being arranged at a cold start. Start or cool heat exchanger 20 is preferably used.
  • warm gas 17A leaves the heat exchanger 20 from the heat exchanger 20.
  • the hot working gas 17A is heated in a further process by a burner 40 and supplied to the expander group 13 as heated working gas 17.
  • the heating of the warm working gas 17A is carried out in this embodiment by means of the combustion of a fuel 41.
  • an electric heater or a second heat exchanger for heating the warm working gas 17A alternatively or cumulatively possible.
  • a burner 40 operated with fuel 41 does not necessarily lead to losses in the efficiency of the internal combustion engine 10, since the expansion of the heated working gas in the expander group 13 also converts the fuel energy converted in the burner 40 into mechanical work. Also conceivable is a combination of a burner 40 with an electric heater or another heat exchanger.
  • FIG. 9 shows a cross section through a cylinder head 50 of an internal combustion engine, in particular also of an internal combustion engine 10 according to the figures explained above.
  • This cylinder head 50 which contains an inlet channel 58 and an outlet channel 59, has a poppet valve 51 in which a pressure compensation is effected by means of a compensating piston 55, whereby a pressure caused by the working gas in the inlet channel 58 is compensated for on the plate 52 of the poppet valve 51 ,
  • the pressure generated in the inlet passage 58 and acting on the plate 52 would, in the event of a positive pressure differential between the inlet passage 58 and the lower pressure in the cylinder, open the poppet valve 51 and allow the working gas in the inlet passage 58 to escape into the cylinder , Since in a conventional internal combustion engine the cylinder internal pressure is always higher than or equal to the pressure in the inlet channel 58, an increased sealing effect on the plate 52 of the poppet valve 51, so that this is not critical in a conventional internal combustion engine.
  • the compensation piston 55 has a piston surface 56, which largely corresponds to the upper plate surface 52A and thus compensates the pressure acting on the upper plate surface 52A accordingly by an opposing force. If only the ambient pressure is applied to the rear-side piston surface 57, the plate valve 51 remains on the lower plate surface 52B. kende cylinder internal pressure.
  • the valve spring 54 can be designed according to common design rules, so that the valve spring 54 only has to apply forces for reliable closing of the poppet valve 51. This design of the spring 54 also has the effect that the cam force necessary for opening the poppet valve 51 is minimal and, as a result, the frictional power of the internal combustion engine 10 is also minimal and the mechanical efficiency of the internal combustion engine 10 is maximal.
  • the piston surface 56 of the balance piston 55 does not necessarily correspond to the upper plate surface 52A. It is also conceivable that the outer diameter of the piston 55 is smaller than the diameter of the plate 52, so that the gas force acting on the balance piston 55 is less than the gas force acting on the upper plate surface 52A. Conversely, the diameter of the compensating piston 55 may be larger than the diameter of the plate 52 of the poppet valve 51. This results in a larger gas force acting on the balance piston 55. Finally, it is also possible to pressurize the rear piston surface 57 with a pressure which corresponds to the cylinder internal pressure. This measure can take place, for example, by means of a connecting channel between the rear piston surface 57 and the expansion cylinder 14.
  • FIG. 9 further shows that the exhaust valve 61 located in the exhaust passage 59.
  • the distinction between the disk 52 of the intake valve 51 and the disk 62 of the exhaust valve 61 can be clearly seen.
  • the diameter of the plate 62 is substantially larger than the diameter of the plate 52.
  • the working gas introduced into the cylinder through the inlet channel 58 is under high pressure in this phase of the process, resulting in a relatively low specific volume and a low volume flow leads. After the expansion of the working gas carried out in the cylinder, the volume of this increases by a multiple, resulting in a considerably larger volume flow in the outlet channel 59. This fact is taken into account by the different plate diameters of the inlet valve 51 and the outlet valve 61.
  • an injection nozzle 63 which allows internal mixture formation by direct injection.
  • the use of an internal mixture formation offers certain degrees of freedom in the injection time and in the injection duration, which can not necessarily be represented by an external mixture formation, since in an external mixture formation, the introduction of the fuel into the cylinder depends only on the opening time and the opening time of the intake valve. Likewise, in an external mixture formation, the regulation of the internal combustion engine 10 by the quantity control is not possible or only to a very limited extent.
  • the exhaust valve 62 is closed approximately between 10 ° and 12 ° before top dead center, so that from this point on the injection of the Fuel can take place, even if the inlet valve 51 has not yet been opened.
  • the indication of the timing of the valves naturally refers to the position of the crankshaft, wherein the top dead center denotes the point at which the piston of the respective expansion cylinder has reached its minimum distance to the combustion chamber roof 60.
  • a non-zero angle offset may be present.
  • the combustion in turn can by numerous measures, such as by the residual gas content, by the temperature of the heated working gas 17, by the degree of further compression in the expansion cylinder 14, by the temperature of the fuel 41 and by the proportion of homogeneous and heterogeneous or external and internal mixture formation are influenced and kept stable.
  • another injection valve can be provided in the inlet channel 58 or even in the cold stream 21 of the heat exchanger 20. be present for the purpose of external mixture formation, which introduces so much fuel in the compressed working gas 16 or the heated working gas 17 that the ignition limits of the fuel used are just not reached.
  • the fuel is thus able to form during a further compression in the expansion cylinder in a very short period of time already radicals or other highly reactive reactants and to accelerate the initiated by the direct injection combustion.
  • FIG. 10 shows a schematic view of a heated accumulator injection system for directly injected fuels, in particular also for a combustion engine according to the figures explained above, in which the heating takes place in the intake manifold. Shown is also the heat exchanger 20 and a plurality of expansion cylinders 14 of the expander 13 of the internal combustion engine 10 described above.
  • the compressed working gas 16 introduced into the heat exchanger 20 is heated in the heat exchanger 20 and distributed as heated working gas 17 to the respective expansion cylinders 14 by means of a working gas distributor or several working gas distributors.
  • the working gas distributor or intake manifold of the expander group 13 includes, as already indicated above, in this case the pressure accumulator 43 of the fuel system.
  • the pressure accumulator 43 is also heated by the heated working gas 17 and transfers the heat from the heated working gas 17 to the fuel 41 to be injected.
  • the fuel 41 is carried in the usual way into a fuel tank 45 with the vehicle and via a fuel pump 46 in FIG the fuel line 42 and consequently also transported into the pressure accumulator 43.
  • the injection nozzle 44 brings the fuel 41 directly into the respective expansion cylinder 14, in which the fuel 41 by the very high temperature of the fuel 41 itself and in the expansion cylinder 14 introduced heated working gas 17 immediately decomposes into its constituents, evaporated with extremely small droplet diameter and thus burned in a very short period of time.
  • the fuel nozzles 44 are controlled as usual and to maintain the greatest possible flexibility of an engine control unit 47 via the control line 48.
  • FIG. 10 shows the previously explained burner 40, which sits in the supply line upstream of the working gas distributor or of the intake manifold of the expander group 13. The burner 40 is used to bring the exiting from the heat exchanger 20 warm working gas 17A to the required process temperature, if no sufficient amount of heat is transferred in the heat exchanger 20.
  • the hot exhaust gas 18 used is not available at the required temperature level after a cold start.
  • the burner 40 or an alternative or further burner upstream of the hot stream 23 of the heat exchanger 20 is housed in a hot exhaust gas 18 leading exhaust pipe.
  • an unillustrated exhaust aftertreatment system may also be provided. In this case, it is also expedient to combine a required further burner with the exhaust aftertreatment system, if the exhaust aftertreatment requires a further increase in temperature of the hot exhaust gas 18.
  • the introduced heat in the heat exchanger 20 can be transferred to the cold stream 21, whereby a feedback of a quantity of heat takes place in the cycle of the internal combustion engine 10.
  • FIG. 11 shows a heat exchanger for an internal combustion engine, in particular also for an internal combustion engine according to the figures explained above.
  • the heat exchanger 20 shown in FIG. 11 is designed as a cross-countercurrent heat exchanger.
  • the introduced into this heat exchanger 20 compressed working gas 16 is transferred into the cold stream 21, which is guided by the existing of straight pipes cold flow chamber 22.
  • the cold stream 21 then leaves the heat exchanger 20 as a heated working gas 17.
  • the heat to this heat exchanger 20 heat to the cold flow 21 transmitted hot stream 23 is formed from the introduced into the heat exchanger 20 hot exhaust gas 18.
  • the hot exhaust gas 18 is introduced in this embodiment by three connecting pipes 26 in the heat exchanger 20 and out again by only one connecting pipe 26, which is why this illustrated heat exchanger 20 is particularly advantageous for an expander with three adjacent expansion cylinders, as an exhaust manifold on the cylinder head of Expander Group 13 is omitted.
  • the hot exhaust gas 18 collected in the inner space 29A and transferred to the hot flow 23, after being introduced by an inlet-side inlet.
  • Schlußrohr 26 passed into the hot flow chamber 24.
  • the hot flow chamber 24 is guided by the located in the heat exchanger 20 intermediate walls 31 in the cross-counterflow to the cold flow chamber 22 and the outlet side discharged as a cold exhaust gas 19 into the environment.
  • the illustrated heat exchanger 20 also has two further features.
  • the heat exchanger 20 includes an insulation 30 located between a first enveloping surface 27 and a second enveloping surface 28.
  • the heat exchanger 20 has a plurality of connecting flanges 25 and connecting tubes 26, which are also designed to be as short as possible to avoid harmful heat losses.
  • the heat exchanger 20 In the simplest embodiment illustrated in this FIG. 11, only a certain volume of air is present in the space 29 between the first enveloping surface 27 and the second enveloping surface 28 as insulation. Air, with its particularly low thermal conductivity, already has an excellent thermal insulation property.
  • the intermediate space 29 can also be evacuated, as a result of which the thermal conductivity can be reduced to a fraction of the thermal conductivity of air.
  • the gap 29 can also be flooded with hot exhaust gas 18, as is done in the inner space 29A also.
  • the gap 29 can be flooded with cold exhaust gas 19, if this has a higher temperature than the temperature of the ambient air.
  • the intermediate space 29 may optionally also have an insulation 30, that is to say a solid insulating layer of granules, ceramics, glass wool or rock wool.
  • the second measure for reducing the heat radiation lies in the design of the connecting tubes 26 and the connecting flanges 25 shown here.
  • the connecting tubes 26, to which the connecting flanges 25 are attached have a length dimensioned such that between a connecting flange 25 and the outer envelope Surface 28 a minimum distance is present, which just barely ensures the mountability of the heat exchanger 20.
  • a flange nut not shown here, is inserted into the space between the connection flange 25 and the outer enveloping surface 28 and a screw guided through the connection flange 25 is fastened by means of the flange nut. So- With, the distance between the connection flange 25 and the outer enveloping surface 28 is such that this distance corresponds at most to 2.1 times the flanged nut height.
  • connection flange 25 can also already be welded or soldered to the connection flange 25.
  • the cohesive connection between the connection flange 25 and the flange nut provides an additional captive protection.
  • the distance of the connecting flange 25 to the outer enveloping surface 28 in this case corresponds to a maximum of 1.1 times the Flanschenmutterière.
  • the third conceivable embodiment results from the fact that the nut is formed together with the connection flange 25 or in that the flange 25 already contains a mounting thread.
  • the distance between the connecting flange 25 and the outer enveloping surface 28 can be further reduced, whereby this distance corresponds to a maximum of 0.1 times the thickness of the connecting flange 25.

Landscapes

  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Exhaust-Gas Circulating Devices (AREA)
  • Engine Equipment That Uses Special Cycles (AREA)

Abstract

L'invention concerne un moteur à combustion interne avec une combustion interne intermittente, avec un groupe compresseur constitué d'au moins un cylindre de compression, avec au moins un groupe d'expansion constitué d'au moins deux cylindres d'expansion, et avec un échangeur de chaleur disposé entre le groupe compresseur et le groupe d'expansion, qui présente un grand rendement. Une chambre d'écoulement froid de l'échangeur de chaleur présente un flux de gaz de travail sortant du groupe compresseur et entrant dans le groupe d'expansion, et une chambre d'écoulement chaud de l'échangeur de chaleur présente un flux de gaz d'échappement sortant du groupe d'expansion. Selon l'invention, le moteur à combustion interne présente, en aval de la chambre d'écoulement froid, un distributeur de gaz de travail avec au moins deux sorties, les sorties du distributeur de gaz de travail étant à chaque fois raccordées à une entrée du cylindre d'expansion.
PCT/DE2011/000189 2010-02-26 2011-02-25 Moteur à combustion interne et procédé pour faire fonctionner un moteur à combustion interne Ceased WO2011103868A2 (fr)

Priority Applications (2)

Application Number Priority Date Filing Date Title
CN2011800112317A CN102844541A (zh) 2010-02-26 2011-02-25 内燃机和用于运行内燃机的方法
DE112011100670T DE112011100670A5 (de) 2010-02-26 2011-02-25 Verbrennungsmotor und Verfahren zum Betrieb eines Verbrennungsmotors

Applications Claiming Priority (6)

Application Number Priority Date Filing Date Title
DE102010009606 2010-02-26
DE102010009606.7 2010-02-26
DE102010047112A DE102010047112A1 (de) 2010-02-26 2010-10-01 Verbrennungsmotor und Verfahren zum Betrieb eines Verbrennungsmotors
DE102010047112.7 2010-10-01
DE102010053778 2010-12-08
DE102010053778.0 2010-12-08

Publications (2)

Publication Number Publication Date
WO2011103868A2 true WO2011103868A2 (fr) 2011-09-01
WO2011103868A3 WO2011103868A3 (fr) 2011-11-10

Family

ID=44507286

Family Applications (1)

Application Number Title Priority Date Filing Date
PCT/DE2011/000189 Ceased WO2011103868A2 (fr) 2010-02-26 2011-02-25 Moteur à combustion interne et procédé pour faire fonctionner un moteur à combustion interne

Country Status (3)

Country Link
CN (1) CN102844541A (fr)
DE (1) DE112011100670A5 (fr)
WO (1) WO2011103868A2 (fr)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2024168427A1 (fr) * 2023-02-13 2024-08-22 Socpra Sciences Et Genie S.E.C. Moteur à combustion interne

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP5951513B2 (ja) * 2013-01-21 2016-07-13 日立オートモティブシステムズ株式会社 多気筒内燃機関の可変動弁装置及び該可変動弁装置の制御装置
WO2025145903A1 (fr) * 2024-01-03 2025-07-10 李永明 Moteur à chambre de combustion externe

Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5857436A (en) 1997-09-08 1999-01-12 Thermo Power Corporation Internal combustion engine and method for generating power

Family Cites Families (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE2732315A1 (de) * 1977-07-16 1979-02-01 Werner Vieler Verbrennungsmotor, insbesondere zweitaktmotor
DE4131289C1 (en) * 1991-09-20 1992-11-12 Mercedes-Benz Aktiengesellschaft, 7000 Stuttgart, De IC engine intake air preheater - has pot-shaped burner housing, penetrating suction manifold, with evaporator secured to housing bottom
BRPI0605059A (pt) * 2006-11-07 2008-07-01 Lucio Arlindo Bueno Villela motor de combustão interna e veìculo contendo tal motor
EP2176530B1 (fr) * 2007-08-13 2013-02-20 The Scuderi Group, LLC Soupapes de moteur équilibrées en pression

Patent Citations (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5857436A (en) 1997-09-08 1999-01-12 Thermo Power Corporation Internal combustion engine and method for generating power

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2024168427A1 (fr) * 2023-02-13 2024-08-22 Socpra Sciences Et Genie S.E.C. Moteur à combustion interne

Also Published As

Publication number Publication date
DE112011100670A5 (de) 2013-02-21
CN102844541A (zh) 2012-12-26
WO2011103868A3 (fr) 2011-11-10

Similar Documents

Publication Publication Date Title
EP2525074B1 (fr) Moteur à combustion interne doté d'un dispositif de chauffage de l'air d'aspiration et procédé de fonctionnement d'un tel moteur à combustion interne
EP2456956A2 (fr) Moteur à pistons axiaux, procédé pour faire fonctionner un moteur à piston axiaux et procédé pour réaliser un échangeur thermique d'un moteur à pistons axiaux
AT3753U1 (de) Brennkraftmaschine, insbesondere mit selbstzündung
EP3901435A2 (fr) Moteur à allumage dans une préchambre et procédé de commande d'un tel moteur
EP2657494A1 (fr) Moteur à combustion à allumage commandé avec une section murale séparant au moins deux cylindres
DE102010047112A1 (de) Verbrennungsmotor und Verfahren zum Betrieb eines Verbrennungsmotors
DE102012214967B3 (de) Vier-Zylinder-Reihenmotor mit Teilabschaltung und Verfahren zum Betreiben eines derartigen Vier-Zylinder-Reihenmotors
EP2456955A2 (fr) Moteur à pistons axiaux, procédé pour faire fonctionner un moteur à pistons axiaux et procédé de réalisation d'un échangeur thermique d'un moteur à pistons axiaux
WO2011009452A2 (fr) Moteur à pistons axiaux et procédé pour faire fonctionner un moteur à pistons axiaux
WO2011103868A2 (fr) Moteur à combustion interne et procédé pour faire fonctionner un moteur à combustion interne
WO2014037068A1 (fr) Dispositif soupape, système comprenant ce dernier et son utilisation pour injecter du carburant sous forme gazeuse
EP3048244B1 (fr) Moteur a pistons axiaux
DE2410948C3 (de) Brennkraftmaschinen-Arbeitsverfahren und nach diesem Verfahren arbeitende Brennkraftmaschinenanlage
DE102014218655B4 (de) Verfahren zum Betreiben einer teilabgeschalteten selbstzündenden Brennkraftmaschine und Brennkraftmaschine zur Durchführung eines derartigen Verfahrens
EP2456954A2 (fr) Moteur à pistons axiaux, procédé pour faire fonctionner un moteur à pistons axiaux et procédé pour réaliser un échangeur thermique d'un moteur à pistons axiaux
DE4120167C2 (de) Verfahren zur Umwandlung von Wärmeenergie in mechanische Bewegungsenergie
DE4134404A1 (de) Verfahren zur umwandlung von waermeenergie in mechanische bewegungsenergie sowie vorrichtung zum druchfuehren dieses verfahrens
CH717353A2 (de) Motor mit Vorkammerzündung und Verfahren zum Steuern eines solchen Motors.
DE102013221109B4 (de) Aufgeladene fremdgezündete Brennkraftmaschine mit Direkteinspritzung und Verfahren zum Betreiben einer derartigen Brennkraftmaschine
DE102019116035B4 (de) Verfahren und Vorrichtung zum Betrieb einer Brennkraftmaschine
EP1092851A2 (fr) Moteur à combustion interne et procédé de fonctionnement
DE102014006471B4 (de) Verfahren zum Betrieb einer Brennkraftmaschine
EP2054607A1 (fr) procédé et dispositif de conversion de l'énergie thermique de combustion en énergie mécanique
DE102008037121A1 (de) Verfahren zum Betreiben einer Brennkraftmaschine sowie Brennkraftmaschine
DE814076C (de) Gemischverdichtende Brennkraftmaschine, insbesondere fuer schwer verdampfbare Brennstoffe

Legal Events

Date Code Title Description
WWE Wipo information: entry into national phase

Ref document number: 201180011231.7

Country of ref document: CN

121 Ep: the epo has been informed by wipo that ep was designated in this application

Ref document number: 11720009

Country of ref document: EP

Kind code of ref document: A2

WWE Wipo information: entry into national phase

Ref document number: 1120111006705

Country of ref document: DE

Ref document number: 112011100670

Country of ref document: DE

REG Reference to national code

Ref country code: DE

Ref legal event code: R225

Ref document number: 112011100670

Country of ref document: DE

Effective date: 20130221

122 Ep: pct application non-entry in european phase

Ref document number: 11720009

Country of ref document: EP

Kind code of ref document: A2