EP3779152A1

EP3779152A1 - An internal combustion engine and a method for operating an internal combustion engine

Info

Publication number: EP3779152A1
Application number: EP19191287.2A
Authority: EP
Inventors: Louis SILEGHEM
Original assignee: Universiteit Gent
Current assignee: Universiteit Gent
Priority date: 2019-08-12
Filing date: 2019-08-12
Publication date: 2021-02-17

Abstract

The invention provides an internal combustion engine (100) comprising a crankshaft (110) and at least one combustion chamber (120). The at least one combustion chamber (120) comprises at least one inlet port (121) for a fuel mixture, and at least one exhaust port (122) for exhaust gases. The at least one combustion chamber (120) comprises a first piston (130) coupled to the crankshaft (110) for reciprocating the first piston (130) in the at least one combustion chamber (120) for a first compression of the fuel mixture during a first duration up to a first compression ratio (CR1) below auto-ignition. The at least one combustion chamber (120) comprises a second piston (140) coupled to a drive system (150) for reciprocating the second piston (140) in the at least one combustion chamber (120) for a second compression of the fuel mixture, in addition to the first compression, during a predetermined second duration, shorter than the first duration, at least up to a second compression ratio for auto-ignition.

Description

Technical field

The present invention relates to an internal combustion engine for homogeneous charge compression ignition (HCCI) or low temperature combustion (LTC). The present invention also relates to a method for operating the internal combustion engine.

Background art

An internal combustion engine for homogeneous charge compression ignition (HCCI) or low temperature combustion (LTC), known from the prior art, such as for example shown in WO 2007/084242 A1 , comprises a crankshaft and at least one combustion chamber or cylinder. The at least one combustion chamber comprises at least one inlet port for a substantially homogenous fuel mixture. The at least one combustion chamber comprises at least one exhaust port for exhaust gases. The at least one combustion chamber comprises a piston. The piston is coupled to the crankshaft for reciprocating the piston in the at least one combustion chamber. The piston is reciprocable for a compression of the fuel mixture up to a compression ratio for auto-ignition of the fuel mixture.
Such an internal combustion engine has the disadvantage that it is difficult to control the timing of the auto-ignition of the fuel mixture over the entire operating range of the internal combustion engine.

Disclosure of the invention

It is an aim of the present invention to provide an internal combustion engine which allows for a better control of the timing of the auto-ignition of the fuel mixture over the entire operating range of the internal combustion engine.
This aim is achieved according to the invention with an internal combustion engine showing the technical characteristics of the first independent claim.
Therefore, the present invention provides an internal combustion engine. Preferably, the internal combustion engine is an internal combustion engine for homogeneous charge compression ignition (HCCI) or low temperature combustion (LTC). The internal combustion engine comprises a crankshaft and at least one combustion chamber or cylinder. The at least one combustion chamber comprises at least one inlet port for a fuel mixture. Preferably, the fuel mixture is a substantially homogenous fuel mixture. Preferably, the fuel mixture is a homogenous fuel mixture. Preferably, the fuel mixture comprises a fuel and an oxidiser. Preferably, the fuel mixture is an air-fuel mixture comprising a fuel and air. The at least one inlet port may comprise an inlet adapted to draw in an air-fuel mixture. Alternatively, the at least one inlet port may comprise an inlet adapted to draw in air and an injection device adapted for injection of a fuel mixture. The at least one combustion chamber comprises at least one exhaust port for exhaust gases. The at least one combustion chamber comprises a first piston. The first piston is coupled to the crankshaft for reciprocating the first piston in the at least one combustion chamber. The first piston is reciprocable for a first compression of the fuel mixture. The first piston is reciprocable for the first compression during a first duration. The first piston is reciprocable for the first compression up to a first compression ratio below auto-ignition of the fuel mixture. The at least one combustion chamber comprises a second piston. The second piston is coupled to a drive system for reciprocating the second piston in the at least one combustion chamber. The second piston is reciprocable for a second compression of the fuel mixture. The second compression is in addition to the first compression. The second piston is reciprocable for the second compression during a predetermined second duration. The second duration is shorter than the first duration. Preferably, the second duration is within the first duration. The second piston is reciprocable for the second compression at least up to a second compression ratio for auto-ignition of the fuel mixture.
The combination of the first piston which provides the first main compression of the fuel mixture during the first duration and the second piston which provides the second compression of the fuel mixture in addition to the first compression for the second duration shorter than the first duration, offers the advantage that the exact timing of the auto-ignition can be controlled more precisely. With internal combustion engines of the state of the art having a single piston providing the compression up to auto-ignition of the fuel mixture the pressure rise due to the compression of said single piston is gradual during the cycle and determined by the crank shaft, such that it is difficult to have exact control of the timing of the auto-ignition of the fuel mixture, for example when the load and/or conditions change. In the internal combustion engine of the present invention, on the other hand, the second piston provides an additional compression and, as such, a pressure rise on top of the pressure rise due to the compression of the first piston. This additional pressure rise due to the compression of the second piston can be made much faster and more precisely timed such that a more precise control of the auto-ignition of the fuel mixture is possible, and this without the first piston having to take care of the full compression of the fuel mixture up to auto-ignition of the fuel mixture.
In the internal combustion engine according to the present invention the fuel mixture may be introduced in the at least one combustion chamber via a single inlet port of the combustion chamber. As an example thereof, the internal combustion engine according to an embodiment of the present invention, may comprise an inlet adapted for drawing in a fuel mixture comprising air and a fuel. Alternatively, the different components of the fuel mixture may be introduced in the at least one combustion chamber via different inlet ports of the combustion chamber for mixing in the combustion chamber. As an example thereof, the combustion engine according to an embodiment of the present invention, may comprise an inlet adapted for drawing in air and an injector adapted for injecting a fuel in the combustion chamber.
It is a further advantage of embodiments of the internal combustion engine according the present invention that the compression ratio may be adjusted during the combustion cycle allowing a more precise control of the timing of the auto-ignition.
In an embodiment of the internal combustion engine according to the present invention the second duration of the second compression is at most 45° crank angle. Preferably, the second duration of the second compression is at most 35° crank angle. More preferably, the second duration of the second compression is at most 25° crank angle. Even more preferably, the second duration of the second compression is at most 15° crank angle.
Having an additional fast and timed pressure rise due to the compression of the second piston offers the advantage of a more precise control of the timing of the auto-ignition of the fuel mixture. Thereby, a trade-off is made between the desired precision in the timing of the auto-ignition of the fuel mixture and the properties of the drive system and the second piston needed for achieving the second compression of the fuel mixture during the second duration up to the second compression ratio.
In an embodiment of the internal combustion engine according to the present invention a top dead centre of the reciprocating motion of the second piston during the second compression is timed within a predetermined range around a top dead centre of the reciprocating motion of the first piston during the first compression.
This embodiment offers the advantage that the first compression of the fuel mixture by the first piston and the second compression of the fuel mixture by the second piston can be combined optimally. In this way the second compression for achieving auto-ignition, which takes place during the short second duration, can be reduced as much as possible.
In an embodiment of the internal combustion engine according to the present invention the predetermined range is at most 10° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. Preferably, the predetermined range is at most 5° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. More preferably, the predetermined range is at most 2° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. Even more preferably, the predetermined range is at most 1° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the internal combustion engine according to the present invention the top dead centre of the reciprocating motion of the second piston during the second compression substantially coincides with the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the internal combustion engine according to the present invention the top dead centre of the reciprocating motion of the second piston during the second compression is timed before or after the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the internal combustion engine according to the present invention the internal combustion engine comprises a control unit. The control unit is operatively connected to the drive system. The control unit is configured for controlling via the drive system at least one characteristic of the reciprocating motion of the second piston. Preferably, the at least one characteristic is selected from the list consisting of the motion profile, the second duration, the second compression ratio and a timing of a top dead centre.
The control unit offers the advantage that the reciprocating motion of the second piston can be adapted to the operating conditions of the internal combustion engine, such as for example the load or combustion behaviour. This allows the desired precision in the timing of the auto-ignition of the fuel mixture to be maintained under the different operating conditions of the internal combustion engine.
In an embodiment of the internal combustion engine according to the present invention the internal combustion engine comprises at least one sensor. The at least one sensor is configured for measuring at least one parameter indicative of the load or combustion behaviour of the internal combustion engine. The control unit is operatively connected to the at least one sensor for retrieving the measured at least one parameter. The control unit is configured for controlling the at least one characteristic of the reciprocating motion of the second piston based on the measured at least one parameter.
The at least one sensor offers the advantage that the control unit can be provided with feedback about the operating conditions of the internal combustion engine, such as for example the load or combustion behaviour. This feedback can then be used by the control unit for adapting the reciprocating motion of the second piston accordingly for controlling the timing of the auto-ignition of the fuel mixture with the desired precision. This embodiment also offers the advantage that the timing of the auto-ignition of the fuel mixture can be adapted each cycle to the load or combustion behaviour for improving performance.
The at least one parameter can for example be the indicated mean effective pressure (IMEP) measured by a pressure sensor. The at least one parameter can for example be a combustion rate, such as for example the crank angle duration after which 50% of the fuel mixture has been burnt (CA50), the crank angle duration after which 10% of the fuel mixture has been burnt (CA10), the crank angle duration over which 10%-90% of the fuel mixture has been burnt (CA10-90), and the pressure rise rate, which parameters can for example be measured by means of a pressure sensor. The at least one parameter can for example be a pressure measured by means of manifold absolute pressure sensor measured at an inlet of the at least one combustion chamber. The at least one parameter can for example be a heat flux measured in the at least one combustion chamber by means of a heat flux sensor.
Furthermore, the present invention also provides a method for operating an internal combustion engine. Preferably, the internal combustion engine is an internal combustion engine for homogeneous charge compression ignition (HCCI) or low temperature combustion (LTC). The method comprises the step of introducing a fuel mixture into at least one combustion chamber or cylinder of the internal combustion engine. Preferably, the fuel mixture is a substantially homogenous fuel mixture. Preferably, the fuel mixture is a homogenous fuel mixture. Preferably, the fuel mixture comprises a fuel and an oxidiser. Preferably, the fuel mixture is an air-fuel mixture comprising a fuel and air. The substantially homogeneous fuel mixture is introduced via at least one inlet port of the at least one combustion chamber. The method comprises the step of reciprocating a first piston of the at least one combustion chamber in the at least one combustion chamber. The first piston is reciprocated by means of a crankshaft of the internal combustion engine. The first piston is reciprocated for a first compression of the fuel mixture. The first piston is reciprocated for the first compression during a first duration. The first piston is reciprocated for the first compression up to a first compression ratio below auto-ignition of the fuel mixture. The method comprises the step of reciprocating a second piston of the at least one combustion chamber in the at least one combustion chamber. The second piston is reciprocated by means of a drive system. The second piston is reciprocated for a second compression of the fuel mixture. The second compression is in addition to the first compression. The second piston is reciprocated for the second compression during a predetermined second duration. The second duration is shorter than the first duration. Preferably, the second duration is within the first duration. The second piston is reciprocated for the second compression at least up to a second compression ratio for auto-ignition of the fuel mixture. The method comprises the step of removing exhaust gases from the at least one combustion chamber via at least one exhaust port of the at least one combustion chamber.
In an embodiment of the method according to the present invention the second duration of the second compression is at most 45° crank angle. Preferably, the second duration of the second compression is at most 35° crank angle. More preferably, the second duration of the second compression is at most 25° crank angle. Even more preferably, the second duration of the second compression is at most 15° crank angle.
In an embodiment of the method according to the present invention a top dead centre of the reciprocating motion of the second piston during the second compression is timed within a predetermined range around a top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the method according to the present invention the predetermined range is at most 10° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. Preferably, the predetermined range is at most 5° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. More preferably, the predetermined range is at most 2° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression. Even more preferably, the predetermined range is at most 1° crank angle from the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the method according to the present invention the top dead centre of the reciprocating motion of the second piston during the second compression substantially coincides with the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the method according to the present invention the top dead centre of the reciprocating motion of the second piston during the second compression is timed before or after the top dead centre of the reciprocating motion of the first piston during the first compression.
In an embodiment of the method according to the present invention the method further comprises the step of a control unit of the internal combustion engine controlling via the drive system at least one characteristic of the reciprocating motion of the second piston. Preferably, the at least one characteristic is selected from the list consisting of the motion profile, the second duration, the second compression ratio and a timing of a top dead centre.
In an embodiment of the method according to the present invention the method comprises the step of measuring at least one parameter indicative of the load or combustion behaviour of the internal combustion engine by means of at least one sensor of the internal combustion engine. The method comprises the step of the control unit retrieving the at least one measured parameter from the at least one sensor. The method comprises the step of the control unit controlling the at least one characteristic of the reciprocating motion of the second piston based on the measured at least one parameter.

Brief description of the drawings

The invention will be further elucidated by means of the following description and the appended figures.

Figure 1 shows a schematic representation of a first embodiment of an internal combustion engine according the present invention.
Figure 2 shows a schematic representation of a second embodiment of an internal combustion engine according the present invention.
Figure 3 shows a schematic representation of a third embodiment of an internal combustion engine according the present invention.
Figure 4 shows a schematic representation of a fourth embodiment of an internal combustion engine according the present invention.
Figure 5 shows the volume displacement of a second piston.
Figure 6 shows the pressure in a combustion chamber with and without the second piston.
Figure 7 shows auto-ignition timings of a methanol-air fuel mixture without the second piston, where the compression ratio is 19.
Figure 8 shows auto-ignition timings of a methanol-air fuel mixture with the second piston, where the first compression ratio is 10 and the second compression ratio is 20.
Figure 9 shows auto-ignition timings of a methanol-air fuel mixture with the second piston, where the first compression ratio is 12 and the second compression ratio is 20.
Figure 10 shows auto-ignition timings of a PRF95 fuel mixture without the second piston, where the compression ratio is 17.
Figure 11 shows auto-ignition timings of a PRF95 fuel mixture with the second piston, where the first compression ratio is 8 and the second compression ratio is 21.
Figure 12 shows auto-ignition timings of a PRF95 fuel mixture with the second piston, where the first compression ratio is 10 and the second compression ratio is 20.
Figure 13 shows the difference in auto-ignition timing of a methanol-air fuel mixture for a range of first and second compression ratios at a first operating point where P₀ is 0.4 bar and T₀ is 288 K and at a second operating point where P₀ is 0.8 bar and T₀ is 308 K.
Figure 14 shows for a methanol-air fuel mixture the effect of the duration of the second compression and the effect of a phase change of the second compression with respect to the first compression.

Modes for carrying out the invention

The present invention will be described with respect to particular embodiments and with reference to certain drawings but the invention is not limited thereto but only by the claims. The drawings described are only schematic and are non-limiting. In the drawings, the size of some of the elements may be exaggerated and not drawn on scale for illustrative purposes. The dimensions and the relative dimensions do not necessarily correspond to actual reductions to practice of the invention.
Furthermore, the terms first, second, third and the like in the description and in the claims, are used for distinguishing between similar elements and not necessarily for describing a sequential or chronological order. The terms are interchangeable under appropriate circumstances and the embodiments of the invention can operate in other sequences than described or illustrated herein.
Moreover, the terms top, bottom, over, under and the like in the description and the claims are used for descriptive purposes and not necessarily for describing relative positions. The terms so used are interchangeable under appropriate circumstances and the embodiments of the invention described herein can operate in other orientations than described or illustrated herein.
The term "comprising", used in the claims, should not be interpreted as being restricted to the means listed thereafter; it does not exclude other elements or steps. It needs to be interpreted as specifying the presence of the stated features, integers, steps or components as referred to, but does not preclude the presence or addition of one or more other features, integers, steps or components, or groups thereof. Thus, the scope of the expression "a device comprising means A and B" should not be limited to devices consisting only of components A and B. It means that with respect to the present invention, the only relevant component.
Within the context of the present invention the term "top dead centre" should be understood as the maximum inwards position of the piston in the at least one combustion chamber during its reciprocating motion. This is the point where the reciprocating motion of the piston reverses in direction from an inwards moving direction to an outwards moving direction.
Within the context of the present invention a duration of X° crank angle should be understood as being expressed relative to the duration T_c of a full rotation of the crank shaft over 360°. Hence, a duration of X° crank angle equals X/360 * T_c.
Figures 1-4 show a schematic representation of an internal combustion engine 100 according to different embodiments of the present invention. The internal combustion engine 100 comprises at least one combustion chamber 120 or cylinder 120, of which only one is shown in the Figures. The combustion chamber 120 is provided with an inlet port 121 for introducing a fuel mixture, preferably a substantially homogeneous fuel mixture, into the combustion chamber 120. The at least one inlet port may comprise an inlet 121 for drawing in air together with an injector 160 for injecting a fuel as is shown in Figure 1. The at least one inlet port may comprise an inlet 121 for an air-fuel mixture as is shown in Figure 2. Such inlet port may comprise an injector 160 for injecting the air-fuel mixture as is shown in Figure 2. The combustion chamber 120 is also provided with an exhaust port 122 for removing exhaust gases out of the combustion chamber 120 after ignition of the fuel mixture in the combustion chamber 120.
The combustion chamber 120 is also provided with a first piston 130. The first piston 130 is coupled to a crankshaft 110 of the internal combustion engine 100 for reciprocating the first piston 130 in the combustion chamber 120. The first piston 130 is reciprocated for a first compression of the fuel mixture inside the combustion chamber 120, such as illustrated in Figure 6. Thereby, the fuel mixture is compressed inside the combustion chamber 120 during a first duration up to a first compression ratio CR1 below auto-ignition of the fuel mixture. Hence, the first compression in itself is insufficient for auto-ignition of the fuel mixture.
Besides the first piston 130, the combustion chamber 120 is also provided with a second piston 140. The second piston 140 is coupled to a drive system 150 for reciprocating the second piston 140 in the combustion chamber 120. The second piston 140 is reciprocated for a second compression of the fuel mixture inside the combustion chamber 120 in addition to the first compression of the fuel mixture by means of the first piston 130, such as illustrated in Figures 5 and 6. Thereby, the fuel mixture is further compressed inside the combustion chamber 120 during a second duration up to a second compression ratio CR2 for auto-ignition of the fuel mixture. Hence, the second compression in addition to the first compression is sufficient for auto-ignition of the fuel mixture.
As such the first piston 130 takes care of a large part of the compression of the fuel mixture over a longer duration, whereas the second piston 140 takes care of the further compression of the fuel mixture over a shorter duration to achieve auto-ignition of the fuel mixture. Thereby, the short pressure rise due to the compression of the second piston 140 enables a more precise timing of the auto-ignition of the fuel mixture than would be possible with a slow gradual pressure rise due to the compression of a single piston.
The second piston 140 may be driven by the drive system 150 such that the top dead centre 141 of the reciprocating motion of the second piston 140 during the second compression substantially coincides with the top dead centre 131 of the reciprocating motion of the first piston 130 during the first compression, for combining the maximum compression of the first piston 130 and the second piston 140 at the same time. Alternatively, the second piston 140 may be driven by the drive system 150 such that the top dead centre 141 of the reciprocating motion of the second piston 140 during the second compression is timed before or after the top dead centre 131 of the reciprocating motion of the first piston 130 during the first compression, but within a predetermined range from the top dead centre 131 of the reciprocating motion of the first piston 130 during the first compression.
The second piston 140 may be arranged opposite of the first piston 130 inside the combustion chamber 120, such as shown in the Figures 1-4. It should however be clear to the skilled person that other arrangements of the position of the second piston 140 with respect to the first piston 130 is also possible. The second piston 140 may for example also be arranged perpendicular to the first piston 130, or under a certain angle with respect to the first piston 130.
The second piston 140 may be arranged such that the second piston 140, similar to the first piston 130, occupies the full bore of the combustion chamber or cylinder 120, such as shown in the Figures 1-3. In alternative embodiments a smaller second piston 140 may however also be used, such as for example shown in Figure 4. In this embodiment, the second piston 140 is arranged in a recess 124 of the combustion chamber 120 which has a smaller bore than the bore of the main part 123 of the combustion chamber 120.
The drive system 150 of the second piston 140 can be designed in different ways, such as shown in the different embodiments of Figures 1-3. As shown in Figure 1, the reciprocating motion of the second piston 140 may for example be driven by means of a cam 152 mounted on a camshaft 151. This camshaft 151 is driven by the crankshaft 110 of the internal combustion engine 100, which is coupled to the camshaft 151 for this purpose by means of a gear system 153. As shown in Figure 2, the reciprocating motion of the second piston 140 may also be driven by means of a spring and catch mechanism 155. As shown in Figure 3, the reciprocating motion of the second piston 140 may also be driven by means of an actuator 156, which may for example be a pneumatic or hydraulic actuator 156. It should however be clear that any other type of suitable drive system 150 known to the skilled person can be used.
The internal combustion engine 100 is also provided with a control unit 170 that is operatively connected to the drive system 150 for controlling the reciprocating motion of the second piston 140. Thereby, characteristics of the reciprocating motion of the second piston 140 such as the second duration, the second compression ratio CR2, and the timing of the top dead centre 141 of the second piston 140 may for example be controlled by means of the control unit 170 via the drive system 150. The use of the control unit 170 givers further control of the exact timing of the auto-ignition of the fuel mixture over the entire operating range of the internal combustion engine 100, because the timing of the auto-ignition of the fuel mixture may for example shift under different operating conditions of the internal combustion engine 100, which can then be corrected for by means of the control unit 170.
The control unit 170 may control the reciprocating motion of the second piston 140 based on the feedback from a sensor 180 in the combustion chamber 120, as shown in Figures 1-4. This sensor 180 is configured for measuring one or more parameters that relate to the load or the combustion behaviour of the internal combustion engine 100. The sensor 180 may for example be a pressure sensor in the combustion chamber 120 or at the inlet port 121, a heat flux sensor, a temperature sensor, a mass flow sensor for the fuel or oxidiser in the fuel mixture or a sensor for measuring ions in the combustion chamber 120, but any other type of suitable sensor known to the skilled person for measuring the load or combustion behaviour of the internal combustion engine 100 can be used. The sensor 180 does not necessarily have to be located in the combustion chamber 120. Other types of sensors may for example be used in other locations, such as for example a sensor for measuring the rotational speed of the crankshaft 110 of the internal combustion engine 120, a sensor for measuring the position of a gas pedal in a vehicle using the internal combustion engine 120, or a sensor for measuring the position of a throttle if a throttle is used.
In an evaluation of the internal combustion engine 100 according to the present invention, the timing of the auto-ignition of a fuel mixture was determined with only a first compression and with a second compression in addition to the first compression, and this over a certain load and temperature range. Starting from a certain pressure P₀ and a temperature T₀, the pressure and temperature during the compression and expansion stroke of the first piston 130 were determined. The initial pressure P₀ was varied between 0.1 and 1.2 bar to simulate different loads, and the initial temperature T₀ was varied between 283 K and 323 K to simulate different inlet air temperatures. As shown in Figure 5, it was assumed that the volume of the combustion chamber 120 decreases and increases following a sine function that could be scaled with the duration and stroke of the second piston 140.
In first instance, the top dead centre 141 of the reciprocating motion of the second piston 140 was chosen to coincide with the top dead centre 131 of the reciprocating motion of the first piston 130. Further below however, with reference to Figure 14, the effect of a phase change, i.e. a change in the timing of the top dead centre 141 of the reciprocating motion of the second piston 140 with respect to the top dead centre 131 of the reciprocating motion of the first piston 130, is also demonstrated.
The duration of the second compression is set to 15° crank angle, of which 7,5° crank angle during the compression stroke and 7,5° crank angle during the expansion stroke. Shorter durations would increase the precision of the timing of auto-ignition of the fuel mixture, but would also increase the pressure rise rate. A phase change of the second compression with respect to the first compression could be used to influence the pressure rise rate. The effect of a change in duration of the second compression on the timing of the auto-ignition of the fuel mixture is also demonstrated.
The stroke of the second piston 140 depends on the compression ratio needed to auto-ignite the fuel mixture. In this application, the compression ratio of the first piston 130 without the second compression is referred to as CR1, and the compression ratio of the second piston 140 in addition to the first compression is referred to as CR2.
An example of the change in the pressure in the combustion chamber 120 due to the reciprocating motion of the second piston 140 is shown in Figure 6. As can be seen on the Figure, this results in a short burst of pressure to ignite the fuel mixture. In this example, CR1 is equal to 12, CR2 is chosen to be 20 and P₀ is equal to 1.2 bar.
First, the auto-ignition timings of a methanol-air fuel mixture with and without the second compression are compared. Thereby, the auto-ignition timings are determined for a load and temperature range at 1500 rpm. Without the second compression, the compression ratio was increased until auto-ignition occurred for the full load and temperature range (P₀ ranging from 0.1 to 1.2 bar and T₀ ranging from 283 to 323 K). To evaluate the auto-ignition timing of the internal combustion engine 100 according to the present invention, CR1 was fixed and CR2 was increased until auto-ignition occurred for every operating point. The results are shown in Figures 7-9.
In Figure 7, the auto-ignition timings of a HCCI internal combustion engine, known from the prior art, are shown. From this Figure, it becomes clear why it is challenging to operate an HCCI engine over a broad load and temperature range. There is a large spread on the auto-ignition timings and for the highest loads and temperatures, the auto-ignition timing is too early which would lead to very high pressure peaks if a big part of the fuel mixture would be burned before top dead centre resulting in low efficiencies, high emissions and even damage to the engine. To avoid this, the compression ratio could be lowered, but this would make it impossible to run the engine in HCCI mode at lower loads without running into problems.
In Figures 8 and 9, the auto-ignition timings of an internal combustion engine 100 according to the present invention are shown for a compression ratio CR1 of respectively 10 and 12. Here, the auto-ignition timings are much more controlled and closer to the top dead centre 131 for the whole load and temperature range.
Next, the same procedure was used to compare the auto-ignition timings for a two-stage auto-ignition fuel mixture (PRF95) with and without the second compression. The results are shown in Figures 8-10.
Without the second compression, as shown in Figure 10, the compression ratio needs to be raised to 17 in order for auto-ignition to occur for the full load and temperature range. Compared to the results on the methanol-air fuel mixture with the second compression, as shown in Figure 7, a big difference in behaviour for the PRF95 fuel mixture can be seen between the low and high loads (P₀ low vs. P₀ high). This can be explained by the two-stage auto-ignition behaviour. At higher loads, the temperatures are already high enough at a point during the compression stroke to trigger the cool-flame reactions. This will promote the main auto-ignition resulting in very early auto-ignition timings during the combustion cycle, more than 25° crank angle before top dead centre 131. At low loads, these cool-flame reactions are not or less triggered, resulting in a more single stage behaviour such as is the case with the methanol-air fuel mixture. This two-stage behaviour results in a big spread of auto-ignition timings shown in Figure 10 for the load and temperature range. This also shows the difficulty of controlling the combustion of two-stage auto-ignition fuels such as gasoline when running an engine in HCCI mode.
In Figures 11 and 12, the results are shown for an internal combustion engine 100 according to the present invention using the PRF95 fuel mixture. CR1 was fixed and CR2 was increased until auto-ignition occurred for every operating point. In Figure 11, CR1 was fixed at 8 and in Figure 12, CR1 was fixed at 10. As can be seen, due to the slower compression in the beginning of the compression stroke compared to what was needed without the second compression (CR1=17 without the second compression), the cool-flame reactions are delayed for the high load points and the auto-ignition timing for the complete load and temperature range is better controlled around the top dead centre 131. This effect is stronger when CR1 is lower, which is clear when comparing Figures 11 and 12. Compared to the results with the methanol-air fuel mixture the spread is still higher, but the difference with a HCCI internal combustion engine, known from the prior art, without the second compression, is still very significant. This demonstrates that the internal combustion engine 100 according to the present invention can also be used to better control the timing of the auto-ignition of more complex fuels with two-stage ignition behaviour.
The more the auto-ignition of the fuel mixture is depending on the fast second compression of the second piston, the more the timing of the auto-ignition of the fuel mixture is controlled by the second compression, which results in a lower spread for the complete load and temperature range. This means, a lower CR1 and a higher CR2 will result in less spread across the load and temperature range. This is demonstrated by means of a methanol-air fuel mixture at two operating points, wherein at the first operating point P₀ is 0.4 bar and T₀ is 288 K, and wherein at the second operating point P₀ is 0.8 bar and T₀ is 308 K, for a range of CR1 and CR2 compression ratios. The difference in auto-ignition timing (degrees crank angle) between the two operating points is shown in Figure 13. As can be seen, the spread is lower for lower CR1 and higher CR2 compressions ratios. However, this does not mean that low CR1 and high CR2 compression ratios are necessarily preferred as this has also an effect on the efficiency, friction, pressure rise rates, etc.

Finally, the influence of the duration and a phase change of the second compression is demonstrated. The duration has been increased from 15° crank angle to 45° crank angle and the top dead centre 141 of the reciprocating motion of the second piston 140 during the second compression has been delayed with 5° crank angle with respect to the top dead centre 131 of the reciprocating motion of the first piston 130 during the first compression. The results are shown in Figure 14. In this Figure, you can see the original results on the methanol-air fuel mixture and the auto-ignition timings for the complete load and temperature range with a delayed phase and/or an increased duration. As can be seen, delaying the second compression shifts the auto-ignition timings to a later time in the cycle. In this way, the timing can be controlled even more for the whole load and temperature range. By increasing the duration of the second compression, the spread of the auto-ignition timings becomes larger. As a result, the shorter the second compression can be, the easier it is to control the auto-ignition timing. However, with very short durations, the pressure rise rate can become very high and the accelerations of the second piston 140 are much higher. The duration will thus be dependent on what is mechanically and thermodynamically acceptable.

References

100	internal combustion engine	150	drive system
110	crankshaft	151	camshaft
120	combustion chamber	152	cam
121	inlet port	153	gear system
122	exhaust port	154	spring
123	main part	155	spring and catch mechanism
124	recess	156	actuator
130	first piston	160	injector
131	top dead centre	170	control unit
140	second piston	180	sensor
141	top dead centre

Claims

An internal combustion engine (100), wherein the internal combustion engine (100) comprises a crankshaft (110) and at least one combustion chamber (120), the at least one combustion chamber (120) comprising at least one inlet port (121) for a substantially homogenous fuel mixture, and at least one exhaust port (122) for exhaust gases, characterised in that the at least one combustion chamber (120) comprises a first piston (130) coupled to the crankshaft (110) for reciprocating the first piston (130) in the at least one combustion chamber (120) for a first compression of the fuel mixture during a first duration up to a first compression ratio (CR1) below auto-ignition, and a second piston (140) coupled to a drive system (150) for reciprocating the second piston (140) in the at least one combustion chamber (120) for a second compression of the fuel mixture, in addition to the first compression, during a predetermined second duration, shorter than the first duration, at least up to a second compression ratio for auto-ignition.
The internal combustion engine (100) according to claim 1, wherein the second duration of the second compression is at most 45° crank angle, preferably at most 35° crank angle, more preferably at most 25° crank angle, and even more preferably at most 15° crank angle.
The internal combustion engine (100) according to claim 1 or 2, wherein a top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression is timed within a predetermined range around a top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The internal combustion engine (100) according to claim 3, wherein the predetermined range is at most 10° crank angle, preferably at most 5° crank angle, more preferably at most 2° crank angle, and even more preferably at most 1° crank angle from the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The internal combustion engine (100) according to claim 3 or 4, wherein the top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression substantially coincides with the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression, or wherein the top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression is timed before or after the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The internal combustion engine (100) according to any one of the claims 1-5, wherein the internal combustion engine (100) comprises a control unit (170) operatively connected to the drive system (150), wherein the control unit (170) is configured for controlling via the drive system (150) at least one characteristic of the reciprocating motion of the second piston (140), preferably selected from the list consisting of the motion profile, the second duration, the second compression ratio and a timing of a top dead centre (141).
The internal combustion engine (100) according to claim 6, wherein the internal combustion engine (100) comprises at least one sensor (180) configured for measuring at least one parameter indicative of the load or combustion behaviour of the internal combustion engine (100), wherein the control unit (170) is operatively connected to the at least one sensor (180) for retrieving the measured at least one parameter, wherein the control unit (170) is configured for controlling the at least one characteristic of the reciprocating motion of the second piston (140) based on the measured at least one parameter.
A method for operating an internal combustion engine (100), wherein the method comprises the steps of:
introducing a fuel mixture into at least one combustion chamber (120) of the internal combustion engine (100) via at least one inlet port (121) of the at least one combustion chamber (120);

reciprocating a first piston (130) of the at least one combustion chamber (120) in the at least one combustion chamber (120) by means of a crankshaft (110) for a first compression of the fuel mixture during a first duration up to a first compression ratio (CR1) below auto-ignition;

reciprocating a second piston (140) of the at least one combustion chamber (120) in the at least one combustion chamber (120) by means of a drive system (150) for a second compression of the fuel mixture, in addition to the first compression, during a predetermined second duration, shorter than the first duration, at least up to a second compression ratio for auto-ignition; and

removing exhaust gases from the at least one combustion chamber (120) via at least one exhaust port (122) of the at least one combustion chamber (120).
The method according to claim 8, wherein the second duration of the second compression is at most 45° crank angle, preferably at most 35° crank angle, more preferably at most 25° crank angle, and even more preferably at most 15° crank angle.
The method according to claim 8 or 9, wherein a top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression is timed within a predetermined range around a top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The method according to claim 10, wherein the predetermined range is at most 10° crank angle, preferably at most 5° crank angle, more preferably at most 2° crank angle, and even more preferably at most 1° crank angle from the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The method according to claim 10 or 11, wherein the top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression substantially coincides with the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression, or wherein the top dead centre (141) of the reciprocating motion of the second piston (140) during the second compression is timed before or after the top dead centre (131) of the reciprocating motion of the first piston (130) during the first compression.
The method according to any one of the claims 8-12, wherein the method further comprises the step of a control unit (170) of the internal combustion engine (100) controlling via the drive system (150) at least one characteristic of the reciprocating motion of the second piston (140), preferably selected from the list consisting of the motion profile, the second duration, the second compression ratio and a timing of a top dead centre (141).
The method according to claim 13, wherein the method further comprises the steps of:
measuring at least one parameter indicative of the load or combustion behaviour of the internal combustion engine (100) by means of at least one sensor (180) of the internal combustion engine (100);

the control unit (170) retrieving the at least one measured parameter from the at least one sensor (180); and

the control unit (170) controlling the at least one characteristic of the reciprocating motion of the second piston (140) based on the measured at least one parameter.