WO2021011977A1

WO2021011977A1 - Method for operating an internal combustion engine

Info

Publication number: WO2021011977A1
Application number: PCT/AT2020/060272
Authority: WO
Inventors: Michael Glensvig; Gernot Graf; Helmut Theissl; Daniele CORSINI
Original assignee: Avl List Gmbh
Priority date: 2019-07-23
Filing date: 2020-07-22
Publication date: 2021-01-28
Also published as: CN114207258A; AT522176B1; AT522176A4

Abstract

The invention relates to a method for operating an internal combustion engine with at least one gas exchange opening per cylinder, with at least one fuel injector, with at least one exhaust gas turbocharger (7) having an exhaust turbine (72) and a compressor (71), wherein exhaust gas heat of an outlet system (3) and/or of an exhaust gas recirculation system (6) of the internal combustion engine (1) is recovered with a waste heat recovery device (5) which operates according to an organic Rankine cycle (OCR). In order to improve the overall thermal efficiency (BTE₁₊₅) there is provision that the internal combustion engine (1) is operated with the following combination of parameters at at least one defined operating point: - compression ratio (CR) between 21 and 23; - peak combustion pressure (P_MX) of at least 250 bar, preferably between 250 bar and 270 bar, in the entire engine characteristic diagram; - volumetric efficiency (λ_ι) of at least 90%; - swirl rate (R_s) in the cylinder between 0 and 1.6; - flow coefficient (K_v) of at least one gas exchange opening: at least 0.068; - maximum injection pressure of fuel injectors: at least 2500 bar; - specific nozzle throughput rate of fuel injectors of 700 - 1200ml/60s per 100 kW rated power, measured at a fuel injection pressure of 100 bar; - maximum exhaust gas turbocharger efficiency: at least 60%, preferably at least 65%; - start of main injection of fuel at the best operating point of thermal efficiency (BTE) of the internal combustion engine (1) so that the centre of combustion (MFB50%) at approximately 6° to 12°, preferably 8° to 10°, crank angle (CA) after the dead centre (ATDC) of the ignition.

Description

Method for operating an internal combustion engine

The invention relates to a method for operating an internal combustion engine with at least one gas exchange opening per cylinder, at least one fuel injector, with at least one exhaust gas turbocharger having an exhaust gas turbine and a compressor, with exhaust gas heat of the exhaust system and / or the exhaust gas recirculation system of the internal combustion engine having an organic rankine -Cycle waste heat recovery device is recovered.

Furthermore, the invention relates to an internal combustion engine with at least one fuel injector, with at least one exhaust gas turbine and a compressor aufwei send exhaust gas turbocharger, the internal combustion engine preferably having an exhaust gas recirculation system with at least one exhaust gas recirculation line between an inlet system and an outlet system, and with one after a Organic Rankine cycle working waste heat recovery device to recover exhaust heat from the exhaust system and / or the exhaust gas recirculation system, which waste heat recovery device has a circuit for a working medium with at least one pump, at least one evaporator, at least one expander and at least one condenser. The invention also relates to a method for designing this internal combustion engine.

Waste heat recovery devices are used, among other things, in commercial road vehicles in order to utilize the waste gas heat from the internal combustion engine using an organic Rankine cycle that has a compressor and an expander. Mechanical work is performed in the expander, for example a turbine or a piston machine.

So far, waste heat recovery devices and internal combustion engines have been developed and optimized independently of one another. The waste heat recovery device was subsequently added to the internal combustion engine. The disadvantage is that - although the internal combustion engine and the waste heat recovery device have been optimized for themselves - the thermal efficiency of the overall system comprising the internal combustion engine and waste heat recovery device was not optimal.

The object of the invention is to improve the overall efficiency of the system internal combustion engine and waste heat recovery device. In particular, it is the object of the invention to achieve an overall thermal efficiency factor of at least 50% in at least one operating point. According to the invention, this object is achieved in that the internal combustion engine is operated in at least one defined operating point with the following combination of parameters:

Compression ratio between 21 and 23;

Peak combustion pressure of at least 250 bar, preferably between 250 bar and 270 bar, in the entire engine map;

Degree of delivery of at least 90%;

Swirl number in the combustion chamber between 0 and 1.6;

Flow coefficient of at least one gas exchange opening: at least 0.068;

Injection pressure of fuel! n per ctors of at least 2500 bar; specific nozzle flow rate of fuel! injectors from 700 - 1200 ml / 60s per 100 kW nominal power, measured at a fuel injection pressure of 100 bar;

Exhaust gas turbocharger efficiency at least 60%, preferably at least 65%;

Start of main fuel injection at the best operating point of thermal efficiency so that the combustion center of gravity (MFB50%) is around 6 ° to 12 °, preferably 8 ° to 10 °, crank angle KW after ignition dead center (ATDC).

If the internal combustion engine has an exhaust gas recirculation system, in particular a high-pressure exhaust gas recirculation system, with which exhaust gas is recirculated from an exhaust system to an intake system of the internal combustion engine in at least one operating point of the internal combustion engine, it is advantageous if the exhaust gas with at least one defined operating point an exhaust gas recirculation rate between 0% and 15% is recirculated. The recirculated exhaust gas is preferably conducted in the exhaust gas recirculation line of the exhaust gas recirculation system via a reed valve which opens in the flow direction.

In one embodiment variant of the invention it is provided that the exhaust gas turbine of the exhaust gas turbocharger is regulated by means of a waste gate or a variable turbine geometry.

In one embodiment of the invention it is provided that the internal combustion engine at a speed of the crankshaft in a range between 1020 U / min and 1200 U / min, preferably between 1025 U / min and 1150 U / min, and at a torque in a range between 75% and 85% of the nominal torque for this speed is operated with maximum thermal efficiency We, the relative boost pressure of the exhaust gas turbocharger is preferably between 2.8 bar and 3.1 bar.

The best results can be achieved if cyclopentane is used as the working medium of the waste heat recovery device.

By the measures mentioned, the mass flow through the internal combustion engine is minimized in order to achieve the highest exhaust gas temperatures. In this way, an overall thermal efficiency of at least 50% is achieved, with the internal combustion engine contributing approximately 48% and the waste heat recovery device approximately 2% to the overall thermal efficiency.

This makes it possible to reduce the C02 emissions compared to conventional combinations of internal combustion engines and waste heat recovery devices of the same power.

In order to achieve a high overall thermal efficiency, a method for designing an internal combustion engine is proposed according to the invention, which has the following steps: a. Adjusting the exhaust gas turbocharger until a minimum boost pressure of the compressor of the exhaust gas turbocharger is reached in which a required amount of exhaust gas can still be recirculated in order to meet the legally prescribed maximum NOx emission values at the tailpipe outlet of the internal combustion engine; b. Increasing the compression ratio of the internal combustion engine until the maximum cylinder peak pressure is reached; c. Repeat steps a. and b. with a higher exhaust gas recirculation rate when the nitrogen oxide content at the tailpipe outlet of the exhaust line of the internal combustion engine exceeds the legally prescribed maximum NOx emission values.

This iterative design method for the internal combustion engine enables the highest overall thermal efficiency for the internal combustion engine including the heat recovery device for a single stationary operating point.

The invention is explained in more detail below using the exemplary embodiment shown in the non-restrictive figures. They show schematically:

1 shows an internal combustion engine according to the invention; Fig. 2 is a waste heat recovery device of this internal combustion engine;

3 shows a detail of the internal combustion engine;

4 shows a simplified waste heat recovery device;

Fig. 5 is a performance diagram of the expander;

6 shows a map of the internal combustion engine without waste heat recovery device; and

Fig. 7 is a map of the internal combustion engine with waste heat recovery device.

Fig. 1 shows schematically an internal combustion engine 1 according to the invention for implementing the claimed method. The internal combustion engine 1 has an inlet line 2, an outlet line 3, an exhaust gas aftertreatment device 4 arranged in the outlet line 3 and a waste heat recovery device 5, of which, in FIG. 1, the exhaust gas heat exchanger 51 and the EGR heat exchanger 52 (EGR = exhaust gas recirculation ) is shown. An exhaust gas recirculation device 6, for example a high-pressure exhaust gas recirculation device, is provided between the outlet line 3 and the inlet line 2, with which exhaust gases are returned from the outlet line 3 to the inlet line 2. Furthermore, the internal combustion engine 1 has an exhaust gas turbocharger 7 with a compressor 71 in the inlet line 2 and an exhaust gas turbine 72 in the outlet line 3.

Gas exchange openings controlled via lift valves, that is to say one or more inlet openings and one or more outlet openings, are provided per cylinder, which enable gas exchange in the cylinder. The internal combustion engine 1 can be self-igniting or externally ignited and can have one or more cylinders for reciprocating pistons.

The exhaust gas aftertreatment device 4 can have at least one particle filter and / or at least one catalytic converter. The exhaust gas recirculation device 6 has an exhaust gas recirculation valve 60 and an exhaust gas recirculation line 61, in wel chem the EGR heat exchanger 52 and a reed valve 62 (flutter valve) is net angeord.

The waste heat recovery device 5 works according to the ORC method (ORC = Organic Rankine Cycle) and has a circuit 50 for an organic working medium, which is shown in FIG. In the circuit 50, in addition to the first evaporator 510 of the exhaust gas heat exchanger 51 and the second Ver evaporator 520 of the EGR heat exchanger 52, an expander 53 - for example one Piston engine or a turbine, a condenser 54, a reservoir 55, a pump 56 and a distributor valve 57 are arranged.

The heat sources - exhaust gas from exhaust line 3 and recirculated exhaust gas from exhaust gas recirculation line 61 - are used in waste heat recovery device 5 to vaporize the working medium in first evaporator 51 and / or second evaporator 52. The first evaporator 51 and the second evaporator 52 are connected in parallel in the circuit 50 of the waste heat recovery device 5 and are connected via the distributor valve 57 to enable operation with or without exhaust gas recirculation. In the latter case, the working medium is directed past the second evaporator 520.

The first evaporator 510 is positioned in the exhaust line 3 after the exhaust gas aftertreatment device 4 of the internal combustion engine 1, as shown in FIG. 3. Because of this, adverse effects on the exhaust gas aftertreatment can be avoided.

By means of the ORC process, the waste heat recovery device 5 converts heat from the exhaust gas (enthalpy) into mechanical energy.

Fig. 4 shows a waste heat recovery device 5 with simplified ORC with only one heat source, namely a first evaporator 510 of an exhaust gas heat exchanger 51. The enthalpy Q of the exhaust gas can flow from the exhaust gas mass flow m, the specific heat capacity c _p , which seeks to be approximately constant and the temperature difference DT on the exhaust gas side between inlet 511 and outlet 512 of the first evaporator 510 can be calculated:

Q = m - c _P - AT

An increase in enthalpy Q results in an increase in mechanical power P of the expander.

In Fig. 5, the power P of the expander 53 is plotted as a function of various ex gas mass flows m and exhaust gas-side inlet temperatures TE of the exhaust gas heat exchanger 51 with a constant efficiency. It can be seen from this that the exhaust-gas-side inlet temperatures T _{E of} the exhaust gas heat exchanger 51 have a greater influence on the power P of the expander 53 than the exhaust gas flow m.

The position of the operating point with the highest thermal efficiency BTEopt for an internal combustion engine according to the invention is shown in the characteristic diagrams shown in FIG. 6 and FIG. 7, the torque MD in each case above the Speed N of the internal combustion engine 1 are plotted. The thermal efficiencies BTEi of the internal combustion engine 1 are entered in the map. Fig. 6 shows the thermal efficiency BTEi for the internal combustion engine 1 alone, i.e. without waste heat recovery device 5 (WHR = Waste Heat Recovery), and Fig. 7 shows the thermal efficiency BTEi ₊ s for the internal combustion engine 1 including waste heat recovery device 5. The marked best point BTE _opt the thermal efficiency BTEi is in Fig. 6 with the internal combustion engine 1 alone at 48% - at a speed N of 1261 rpm and a torque MD of 1892 Nm, that is 84% of the nominal torque MN for this speed N - and in Fig. 7 in the internal combustion engine 1 including waste heat recovery device 5 at about 50% - at a speed N of 1230 rpm and a torque MD of 1789 Nm, which is 80% of the nominal torque MN for this speed N.

On the basis of FIGS. 5, 6 and 7, the conclusion can be drawn that - with a given amount of waste heat from an internal combustion engine 1 - high exhaust gas temperatures TE and low exhaust gas flows m are better than low ones for the efficiency of the ORC process Exhaust gas temperatures TE and high exhaust gas mass flows m.

High exhaust gas temperatures TE with the lowest possible exhaust gas mass flows m can be achieved in an internal combustion engine 1 by reducing the amount of air as much as possible.

Cause and effect relationship for a diesel engine

In the following, two different technological concepts for achieving a high thermal efficiency BTEi for the internal combustion engines 1 alone are shown, with the NOx emission requirements still being met

Concept 1 (low cylinder mass):

• “Low” cylinder mass enables the exhaust gas turbocharger 7 to be designed with “low” boost pressure and positive pumping work.

• The low cylinder mass enables a high compression ratio CR with an engine design for a fixed ignition peak pressure (e.g. 250 bar).

However, the positive pumping work prevents the transport of large quantities of recirculated exhaust gas from the high-pressure exhaust gas recirculation device, which leads to high NOx emissions from internal combustion engine 1. Therefore, an efficient but expensive De-NOx system (for example SCR catalytic converter) of the exhaust gas aftertreatment device 4 is required. Concept 1 has the following effects on the thermal efficiency BTEi of the internal combustion engine 1 and the conditions for the first exhaust gas evaporator 510 of the waste heat recovery device 5:

• Positive pumping work (higher intake pressure than exhaust pressure) increases the thermal efficiency of the internal combustion engine 1.

A high compression ratio CR increases the combustion efficiency of the high pressure cycle and enables a higher thermal efficiency BTEi of the internal combustion engine 1.

• A lower cylinder mass, however, increases the heat transfer to the cylinder walls, which in turn reduces the BTEi thermal efficiency.

The lower air flow and the resulting lower boost pressure lead to a lower pressure in front of the exhaust gas turbine 72 of the exhaust gas turbo charger 7, which in turn leads to a lower expansion in the exhaust gas turbine 72 and a higher exhaust gas temperature TE after the exhaust gas turbine 7 enables.

With the same exhaust gas enthalpy, the exhaust gas temperature TE is "high" and the exhaust gas mass flow m is "low", which is favorable for the thermal efficiency BTEs of the waste heat recovery device 5.

Concept 2 (high cylinder mass):

• High cylinder mass leads to a turbocharger design with high boost pressure and typically negative pumping work

• The high cylinder mass enables a low compression ratio with an engine design for a fixed ignition peak pressure (e.g. 250 bar).

• The negative pumping work enables the transport of large amounts of recirculated exhaust gas from the high-pressure exhaust gas recirculation system, which leads to lower NOx emissions from internal combustion engine 1 and a simpler and more cost-effective De-NOx system (e.g. SCR catalytic converter) for the exhaust gas aftertreatment device 4 allows.

Concept 2 has the following effects on the thermal efficiency BTEi of the internal combustion engine 1 and the conditions for the first exhaust gas evaporator 510 of the waste heat recovery device 5:

• A higher cylinder mass reduces the heat transfer to the cylinder walls, whereby the thermal efficiency BTEi of the internal combustion engine 1 is increased.

A negative pumping work (higher exhaust gas pressure than inlet pressure) reduces the thermal efficiency BTEi of the internal combustion engine 1. A low compression ratio CR due to the PFP design limit (PFP = peak firing pressure) leads to a lower combustion efficiency, which in turn leads to a lower thermal efficiency BTEi of the internal combustion engine 1.

The higher air flow and the resulting higher boost pressure lead to a higher pressure in front of the exhaust gas turbine 72, which in turn leads to a higher expansion in the exhaust gas turbine 72 and enables a lower exhaust gas temperature TE after the exhaust gas turbine 72.

With the same exhaust gas enthalpy, the exhaust gas temperature TE is low and the exhaust gas mass flow m is “high”, which is unfavorable for the thermal efficiency BTEs of the waste heat recovery device 5.

Both for concept 1 and for concept 2, the following functions are used to achieve the highest thermal efficiency BTEi of the internal combustion engine 1:

• The valve control of the gas exchange valves enables a degree of delivery li of> 90% at the best point when no Miller control is available;

• low swirl numbers Rs in the combustion chamber in the range from 0 to 1.6;

• maximum flow coefficient K _{v of} the gas exchange openings: equal to or higher than 0.068;

• Maximum fuel injection pressure of the fuel! injectors equal to or above 2500 bar;

• Nozzle flow rate of the fuel injectors: 700 - 1200 ml / 60s per 100 kW nominal power PWR at a fuel injection pressure of 100 bar;

• Reed valves 62 in the exhaust gas recirculation line 61 for internal combustion engines 1 with exhaust gas recirculation system 6;

• Maximum turbocharger efficiency over 60%;

• Start of the main injection in the best operating point of the thermal efficiency BTEi, whereby a combustion center of gravity MFB50% is reached at 6 - 12 ° crank angle KW after the dead center ATDC of the ignition.

Simulations that compare the thermal efficiency BTEi ₊ s of the internal combustion engine 1 together with the waste heat recovery device 5 of Concept 1 and Concept 2 are shown in Table 1 below. Both concepts have a thermal efficiency BTEi of the internal combustion engine 1 alone of 48%. Cyclopentane was used as the working medium of the waste heat recovery device 5.

Concept 1:

Measures to reduce the specific fuel consumption BSFC:

• High compression ratio CR

• Lower boost pressure leads to positive pumping work (P_im> P_31) Result:

• Less EGR transport and higher NOx

• Lower cylinder mass

• Less expansion of the exhaust gas turbine

• Higher exhaust gas temperature

Concept 2:

Measure to reduce the specific fuel consumption BSFC:

• High cylinder mass results in reduced heat transfer Result:

• High boost pressure leads to negative pumping work (PJm <P_31), which enables EGR transport and reduces NOx

• Lower compression ratio with the same peak pressure PFP

• High expansion in the exhaust gas turbine

• Lower exhaust temperature

The results show that the total thermal efficiency BTEi ₊ s for internal combustion engine 1 including waste heat recovery device 5 for concept 1 is much higher than for concept 2. The main reason for this is the higher exhaust gas temperature TE before entering the evaporator 51, which is required for the ORC process is favorable.

Concept 1 thus achieves, in combination with the waste heat recovery device 5 using the ORC process, the highest overall thermal efficiency BTEi ₊ s.

Based on these findings, a design process for achieving the highest thermal efficiency BTEi ₊ s for an internal combustion engine 1 including waste heat recovery device 5 can be found in a single stationary map point:

This is a design method for the highest thermal efficiency BTEi ₊ s in a single map point for an internal combustion engine 1 with high pressure exhaust gas recirculation including a waste heat recovery device 5 for at least one waste heat source, the internal combustion engine 1 having a fixed ignition peak pressure (PFP), a non-variable valve control and a fixed volumetric efficiency, a fixed maximum fuel injection pressure (rail), a fixed nozzle flow rate, at least one reed valve 62, a fixed turbocharger efficiency (e.g. 62%) and a start of the main injection, which has a combustion focus MFB50% at has about 6 ° to 12 ° crank angle KW after the dead center ATDC of the ignition in a single operating point, which meets the legally prescribed maximum NOx emission values at the tailpipe outlet:

The following iteration is carried out:

1) Adaptation of the exhaust gas turbocharger 7, which enables a reduction in the boost pressure to the point at which the required recirculated exhaust gas can still be transported in order to meet the statutory NOx emission regulations. There will be a compromise entered between positive pumping work and cylinder mass. The required boost pressure or the required exhaust gas recirculation rate RT_EGR is dependent on the performance of the exhaust gas treatment device 4.

2) The compression ratio CR is increased until the maximum cylinder peak pressure is reached.

3) The resulting exhaust gas temperature, the exhaust gas mass flow and the nitrogen oxide NOx_Engine_Out emitted by the internal combustion engine 1 enable a nitrogen oxide content NOx_TailPipe on the end pipe 31 for the given exhaust gas aftertreatment device 4.

4) If the nitrogen oxide content NOx_TailPipe on the tailpipe exceeds the legally prescribed maximum NOx emission values at the tailpipe outlet 31, then step 1) is repeated with a higher exhaust gas recirculation rate RT_EGR (iteration).

This iterative design method for the internal combustion engine 1 enables the highest overall thermal efficiency BTEi ₊ s for the internal combustion engine 1 including waste heat recovery device 5 for a single stationary operating point.

Claims

PATENT CLAIMS

1. A method for operating an internal combustion engine with at least one gas exchange opening per cylinder, with at least one fuel! The injector, with at least one exhaust gas turbine (72) and a compressor (71), has the exhaust gas turbocharger (7), with exhaust gas heat of an exhaust system (3) and / or an exhaust gas recirculation system (6) of the internal combustion engine (1) having an organic rankine -Cycle (ORC) working waste heat recovery device (5) is recovered, characterized in that the internal combustion engine (1) is operated in at least one defined operating point with the following combination of parameters:

Compression Ratio (CR) between 21 and 23;

Peak combustion pressure (P_MX) of at least 250 bar, preferably between 250 bar and 270 bar, in the entire engine map;

Degree of delivery (li) of at least 90%;

Swirl number (Rs) in the cylinder between 0 and 1.6;

Flow coefficient (K _v ) of at least one gas exchange opening: Min least 0.068;

maximum injection pressure of fuel! injectors: at least 2500 bar

specific nozzle flow rate of fuel! n per ctoren from 700 - 1200ml / 60s per 100 kW nominal power, measured at a fuel injection pressure of 100 bar;

Maximum exhaust gas turbocharger efficiency: at least 60%, preferably at least 65%;

Start of the main injection of the fuel at the best operating point of the thermal efficiency (BTE) of the internal combustion engine (1) so that the combustion center of gravity (MFB50%) is around 6 ° to 12 °, preferably 8 ° to 10 °, crank angle (KW) after dead center (ATDC) of the ignition.

2. The method according to claim 1, wherein by means of an exhaust gas recirculation system (6) to at least the defined operating point of the internal combustion engine (1) exhaust gas from an exhaust system (3) to an intake system (2) is recirculated, characterized in that the exhaust gas is at least in the defined Operating point with an exhaust gas recirculation rate (RT_EGR) between 0% and 15% is returned.

3. The method according to claim 2, characterized in that the recirculated exhaust gas is passed in an exhaust gas recirculation line (61) of the exhaust gas recirculation system (6) via a reed valve (62) which opens in the flow direction.

4. The method according to any one of claims 1 to 3, characterized in that the exhaust gas turbine (72) of the exhaust gas turbocharger (7) is controlled by means of a waste gate or a variable turbine geometry.

5. The method according to any one of claims 1 to 4, characterized in that the internal combustion engine (1) at a speed (N) in a range between 1020 U / min and 1200 U / min, preferably between 1025 U / min and 1150 U / min, and at a torque (M) in a range between 75% and 85% of the nominal torque (MN) for this speed (N) with maximum thermal efficiency (BTEi), whereby preferably the relative boost pressure of the exhaust gas turbocharger (7 ) is between 2.8 bar and 3.1 bar.

6. The method according to any one of claims 1 to 5, characterized in that cyclopentane is used as the working medium of the waste heat recovery device (5).

7. Internal combustion engine (1) with at least one gas exchange valve, with at least one fuel! injector, with at least one exhaust gas turbine (72) and a compressor (71) having exhaust gas turbocharger (7), preferably as the internal combustion engine (1) an exhaust gas recirculation system (6) with at least one exhaust gas recirculation line (61) between an inlet system (2) and an exhaust system (3), and with a waste heat recovery device (5) operating according to an organic Rankine cycle (ORC) for recovering exhaust gas heat from the exhaust system (3) and / or the exhaust gas recirculation system (6), which waste heat recovery device (5) has a circuit for a working medium with at least one pump (56), at least one evaporator (510, 520), at least one expander (53) and at least one condenser (54), characterized in that the internal combustion engine (1) is designed to operate in at least to be operated at a defined operating point with the following combination of parameters:

Compression Ratio (CR) between 21 and 23;

Degree of delivery (li) of at least 90%;

Swirl number (Rs) in the cylinder between 0 and 1.6; Flow coefficient (K _v ) of at least one gas exchange opening: at least 0.068;

maximum injection pressure of at least one fuel! n per ktors: at least 2500 bar;

specific nozzle flow rate of the at least one fuel! injectors from 700 - 1200ml / 60s per 100 kW nominal power (PWR), measured at a fuel injection pressure of 100 bar;

Exhaust gas recirculation rate (RT_EGR) between 0% and 15%;

Start of the main injection of the fuel at the best operating point of thermal efficiency so that the focus of combustion (MFB50%) is around 6 ° to 12 °, preferably 8 ° to 10 °, crank angle (KW) after ignition dead center (ATDC) .

8. Internal combustion engine (1) according to claim 7, with an exhaust gas recirculation system (6), preferably a high-pressure exhaust gas recirculation system, with at least one exhaust gas recirculation line (61) between an inlet system (2) and an outlet system (3), characterized in that the exhaust gas in the defined operating point with an exhaust gas recirculation rate (RT EGR) between 0% and 15%.

9. Internal combustion engine (1) according to claim 8, characterized in that a reed valve (62) which opens in the flow direction of the recirculated exhaust gas is arranged in at least one exhaust gas recirculation line (61) of the exhaust gas recirculation system (6).

10. Internal combustion engine (1) according to one of claims 7 to 9, characterized in that the exhaust gas turbine (72) of the exhaust gas turbocharger (7) has a waste gate or a variable turbine geometry.

11. Internal combustion engine (1) according to one of claims 7 to 10, characterized in that the maximum thermal efficiency (BTEi) of the internal combustion engine (1) at a speed (N) in a range between 1020 rpm and 1200 rpm, preferably between 1025 rpm and 1150 rpm, and with a torque (M) in a range between 75% and 85% of the nominal torque (MN) for this speed (N), whereby preferably the relative boost pressure of the exhaust gas turbocharger (7 ) is between 2.8 bar and 3.1 bar.

12. Internal combustion engine (1) according to one of claims 7 to 11, characterized in that the working medium of the waste heat recovery device (5) is cyclopentane.

13. A method for designing an internal combustion engine (1) according to one of claims 7 to 12, characterized in that the following steps are carried out:

a. Adjusting the exhaust gas turbocharger (7) until a minimum charge pressure of the compressor (71) of the exhaust gas turbocharger (7) is reached, in which a required amount of exhaust gas can still be recirculated to the legally prescribed maximum NOx emission values at the tailpipe outlet (31) of the internal combustion engine (1) to meet; b. Increasing the compression ratio (CR) of the internal combustion engine (1) until the maximum cylinder peak pressure is reached, c. Repeat steps a. and b. with a higher exhaust gas recirculation rate (RT EGR) when the nitrogen oxide content (NOx_TailPipe) exits the tailpipe (31) of the internal combustion engine (1) exceeds the statutory maximum NOx emission values.

2020 07 22

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