RU158763U1 - INFLATED COMBUSTION ENGINE - Google Patents

Abstract

1. A supercharged internal combustion engine, comprising: - an intake system for supplying the internal combustion engine with charge air, - an exhaust system for exhausting exhaust gases, and - at least one exhaust gas-driven turbocharger (1), in which one and the same the same rotatable shaft (8) has a compressor (2) located in the intake system and an axial turbine (3) located in the exhaust system, the compressor (2) comprising a housing (2a) and at least two rotors (2b , 2b), located in the specified housing (2a), which are mounted on the shaft (8) and are made with the possibility of separately loading the charge air through the intake system by means of supply channels (4b, 4b), which are separated from each other and are made in one piece with housing (2a), characterized in that - the supply channels (4b, 4b) of at least two rotors (2b, 2b) contain inlet channels (5b, 5b), which are located next to with each other along the shaft (8) and run radially relative to the shaft (8). 2. The supercharged internal combustion engine according to claim 1, characterized in that the supply ducts (4b, 4b) are made in such a way that the following is observed: 0.9≤p / p≤1.1, moreover, the robot means the pressure immediately before the first rotor (2b ), and rob means the pressure just before the second rotor (2b). 3. A supercharged internal combustion engine according to claim 1 or 2, characterized in that the supply ducts (4b, 4b) are made in such a way that the following is observed: 0.95≤p / p≤1.05, moreover, the robot means the pressure immediately before the first rotor (2b), and rob means the pressure just before the second rotor (2b). 4. Move

Description

This utility model relates to a supercharged internal combustion engine comprising:

- an intake system for supplying an internal combustion engine with charge air,

- an exhaust system for exhaust gas, and

- at least one turbocharger driven by exhaust gases,

in which a compressor arranged in the intake system and an axial turbine located in the exhaust system are installed on one rotatably shaft, the compressor comprising a housing and at least two rotors that are located in the housing, mounted on the shaft, and can be separately loaded with charge air through the intake system by means of supply channels, which are separated from each other and are made integrally with the housing.

The internal combustion engine of the type described is used, for example, as a traction engine of a car. In the context of this utility model, the expression “internal combustion engine” encompasses diesel engines, Otto cycle engines, but also hybrid internal combustion engines that use a hybrid combustion process, and hybrid drives that include not only an internal combustion engine, but also an electric motor, which in the drive system is connected to the internal combustion engine and receives energy from the internal combustion engine, or which as a switchable auxiliary The leg drive creates extra power output.

State of the art

Internal combustion engines are increasingly equipped with pressurization (see, for example, DE 38 38 149 A1), and pressurization is mainly a way to increase power, in which the pressurized air required for the combustion process in the engine is compressed, resulting in a large mass of charge air can be fed into the working cycle in each cylinder. Thus, the mass of fuel and, consequently, the average pressure can be increased.

For boosting, an exhaust gas driven turbocharger is often used, in which the turbine and compressor are located on the same shaft, and a stream of hot exhaust gas is supplied to the turbine, which expands in the turbine, which releases energy that drives the shaft into rotation. The energy that is transferred to the shaft by the flow of exhaust gas is used to drive a compressor, which is likewise connected to the shaft. The compressor compresses the air supplied to it and delivers it, as a result of which cylinder pressurization is realized.

An advantage of a turbocharger driven by exhaust gases, for example, compared to a mechanical supercharger, is that no mechanical connection is required to transfer power between the supercharger and the internal combustion engine. While a mechanical supercharger draws the energy necessary for its drive directly from the internal combustion engine, and thereby reduces the output power and efficiency, a turbocharger driven by exhaust gases uses the energy of the hot exhaust gases at the exhaust from the engine.

However, a mechanical supercharger has an advantage over an exhaust gas driven turbocharger in that it is possible to obtain a substantially high boost pressure even at a low engine speed, since in the case of a turbocharger, a drop in torque is observed at an insufficiently high shaft speed . The drop in torque can be explained if we take into account that in the case of a turbocharger, the pressure ratio on the boost compressor depends on the pressure ratio on the turbine. For example, if the engine shaft speed decreases in the case of a diesel engine, or if the load decreases in the case of an Otto engine, this leads to a reduced mass flow of exhaust gas, and therefore to a lower pressure ratio on the turbine. As a result, the lower the rotational speed of the engine shaft or, accordingly, the load, the more the pressure ratio on the boost compressor decreases, which is equivalent to a drop in torque.

In this case, in principle, the drop in boost pressure can be compensated by reducing the cross section on the turbine, thereby increasing the pressure ratio on the turbine accordingly. Therefore, in practice, a relatively small turbocharger is often used, that is, a turbocharger with a small cross section of the turbine, in combination with an exhaust gas bypass device. Such a turbine is often referred to as a bypass damper turbine. If the mass flow of exhaust gas exceeds a critical value, then part of the flow of exhaust gas is diverted to bypass the turbine along the bypass line. To ensure sufficient boost even at a high engine speed, and / or to provide a relatively large amount of exhaust gas, and / or to further improve the torque characteristics of a supercharged internal combustion engine, a number of turbochargers are often provided that are connected in parallel or in series, and if desired, in combination with a mechanical supercharger.

The internal combustion engine according to the present utility model also comprises at least one turbocharger driven by the exhaust gas, and the compressor comprises at least two rotors mounted for rotation.

The latter is due to the fact that on the side of the exhaust gas outlet a small turbine is preferable, the small rotor of which provides a quick response due to low inertia, and for the reasons mentioned above, provides a sufficiently high boost pressure at low speeds and / or small amounts of exhaust gas. On the contrary, on the inlet side, the compressor must provide or must be able to provide the maximum possible amounts of charge air, and in fact it is required that the cross section of the compressor be correspondingly large, that is, that the compressor has a relatively large rotor. However, studies show that excessively large differences in rotor diameters spoil performance and reduce the efficiency of a turbocharger driven by exhaust gas. A situation in which the diameter of the compressor rotor is significantly larger than the diameter of the turbine rotor is unfavorable and should be avoided.

For the above reasons, it is advantageous to equip the compressor with several, i.e. at least two small rotors instead of one large rotor. For example, FIG. 1 is a cross-sectional view of a turbocharger 21 driven by an exhaust gas and comprising a compressor 22 of this type.

The compressor 22 and the turbine 23 of the turbocharger 21 have a common shaft 28 mounted for rotation; the turbine 23 has a housing 23a, the turbine rotor 23b being located in said housing 23a; the compressor 22 has a housing 22a and two rotors 22b ₁ , 22b ₂ , which are located in the specified housing 22a and mounted on the shaft 28.

The turbine 23 is made in the form of an axial turbine, i.e. the gas flow running on the rotor blades moves essentially in the axial direction. In the context of the present utility model, “substantially in the axial direction” means that the velocity component in the axial direction is greater than the radial velocity component. In this case, the free-stream velocity vector in the region of the rotor 23b is directed substantially parallel to the shaft 28 of the turbocharger 21.

In this case, passing through the turbine inlet 23a ′, the exhaust gas is initially directed radially by means of a spiral housing 23a ′ that extends completely around, and then the flow is deflected substantially axially towards the rotor 23b. In this case, the flow approaches the rotor 23b with a swirl even without a guiding device, and the swirl is used in the turbine 23 to extract energy, i.e. to form boost pressure. In this case, the exhaust region extends coaxially with the shaft 28 of the axial turbine, as a result of which the exhaust gas in the outlet region does not experience any deviation or change in flow direction.

The exhaust pipe 22a ′ of the compressor 22 is likewise made in the form of a spiral housing 22a ′ that extends completely around and is connected to the radially outwardly directed exhaust channel 27 of the two rotors 22b ₁ and 22b ₂ . Compressor 22 is a radial compressor. In the case of the compressor 22, the term “radial” refers to the direction of the flow exiting from the rotor blades. That is, in the case of a radial compressor, the flow exits substantially radially. In this context, “substantially radially” means that the velocity component in the radial direction is greater than the axial velocity component. The flow velocity vector crosses the shaft 28 of the compressor 22 substantially at right angles if the output stream is exactly radial. The oncoming flow may, and in the preferred case, should be axial.

On the contrary, the power of the two rotors 22b ₁ , 22b ₂ , the rear sides of which are adjacent to each other, i.e. the rear sides of which are directed opposite to each other and touch without a gap, is carried out separately from each other. The first rotor 22b ₁ , located at the end of the shaft 28 closest to the compressor, on the side that is facing away from the turbine 23, axially flows a gas stream through the first supply channel 24b _1, supplying air for boosting, while the second rotor 22b ₂ located between the first the rotor 22b ₁ and the turbine 23, is under the influence of a gas stream moving axially and entering through the second inlet channel 24b ₂ (i.e. is supplied with air to be boosted).

The channel for supplying charge air to the second compressor rotor 22b ₂ through the second supply channel 24b ₂ , which, starting from the input channel 25b ₂ , passes radially with respect to the shaft 28 in the direction of the rotor 22b ₂ , follows curvilinearly and has a section near the rotor end substantially axial supply passage 26b ₂ because the task of ensuring axial flow ramp 22b on the rotor ₂ requires charge air flow deflection with change of direction at a relatively large value, this leads to a loss of pressure in the flow supercharging ary air.

In contrast to the air supply to the second compressor rotor, the charge air supply to the first compressor rotor 22b ₁ occurs in the axial direction. As a result, there is no deviation or change in the direction of flow of charge air, and no pressure drop is observed.

As a result of the fact that the first inlet channel 24b ₁ together with the inlet channel 25b ₁ and the supply channel 26b ₁ from the rotor, right up to the rotor 22b ₁ , are aligned with the shaft 28 of the turbocharger 21, the inflow of charge air is essentially axial. A power supply of this type avoids the relatively large pressure losses associated with the deviation of the flow, which leads to a higher flow rate of charge air, which is characterized by an inlet pressure p _{1, in} , which is higher than the inlet pressure p _{2, in} immediately before the second rotor 22b ₂ .

Since both rotors 22b ₁ , 22b ₂ share a common outlet channel 27, the output pressures p _{1, out} and p _{2, out are the} same in magnitude, but the two rotors 22b ₁ , 22b ₂ have different compressor pressure ratios, respectively, p _{1, in} / p _{1, out} and p _{2, in} / p _{2, out} .

However, in the case of a compressor of this type, the task and preference is to obtain the same ratios of pressures and flows of charge air through individual rotors. In particular, it is necessary to exclude the situation in which one rotor transfers energy to another rotor, that is, the circulation of charge air occurs, and / or the rotor completely stops the transfer of charge air and overheats.

Utility Model Disclosure

In contrast to the aforementioned situation typical of the prior art, the purpose of this utility model is to provide a supercharged internal combustion engine corresponding to the preamble of paragraph 1 of the utility model formula, by which the disadvantages inherent in the prior art, for example, DE 38 38 149A1, and in particular, are eliminated improved supply of compressor rotors with charge air.

The specified goal is achieved due to the supercharged internal combustion engine, comprising:

- an intake system for supplying an internal combustion engine with charge air,

- an exhaust system for exhaust gas, and

- at least one exhaust gas driven turbocharger in which a compressor located in the intake system and an axial turbine located in the exhaust system are installed on the same rotatably shaft, the compressor comprising a housing and at least two rotors located in the specified housing, which are mounted on the shaft, and can be separately loaded with charge air through the intake system by means of supply channels that are separated from each other and made wives as one with the body,

wherein the internal combustion engine differs from the prior art in that

the supply channels of said at least two rotors contain, as sections, input channels that are adjacent to each other along the shaft and extend radially relative to the shaft.

In contrast to the prior art, the internal combustion engine according to the present utility model has similar supply channels that direct charge air to the compressor rotors. Air supply occurs through symmetrical supply channels in the broadest sense of the word. That is, the profiles of the supply channels, starting from the input channel and up to the rotor at the end of the supply channel closest to the rotor, are similar to each other, i.e., are comparable. As a result, the direction changes experienced by the charge air flow passing through the supply channels, and therefore the pressure losses and the inlet pressures p _{1, in} , p _{2, in} , ... p _{i, in} directly in front of the i rotors, are similar to each other, where i denotes the number of rotors.

Since the rotors share a common exhaust channel, the output pressures p _{1, out} , p _{2, out} , ... p _{i, out} are equal in magnitude, and all rotors have equal or close pressure ratios on the compressor p _{1, in} / p _{1, out} , p _{2, in} / p _{2, out} , ... p _{i, in} / p _{i, out} and p _in / p _out , respectively.

Equal pressure ratios and charge air flows through individual rotors significantly improve the performance of the compressor, and, consequently, the exhaust turbocharger. In particular, a situation is excluded in which one rotor transfers charging air to another rotor and there is a circulation of charging air, or some rotor ceases to transmit charging air at all and overheats.

The turbine of an internal combustion engine according to the present utility model is an axial turbine. Axial design facilitates exhaust gas supply. Relatively large pressure losses in the exhaust gas flow resulting from flow deviations are eliminated. As a result, the turbine provides very high energy exhaust gas, namely, an inlet leading to the turbine rotor.

As a result, due to the internal combustion engine according to the present utility model, the objective of the utility model is achieved - to create a supercharged internal combustion engine in which there are no disadvantages inherent in the prior art, and in particular, the supply of compressor rotors with boost air is improved.

Each inlet channel of an internal combustion engine according to the present utility model extends in the same way, starting from the inlet channel through which charge air is supplied, and up to the corresponding rotor, each inlet channel in the form of a section at the end closest to the rotor and contains a supply channel. In order to realize or ensure identical or similar flow ratios in the supply channels, and therefore similar inlet pressures p _{i, in} immediately in front of the rotors, all inlet channels extend radially relative to the shaft, and according to the present utility model, the inlet channels are located next to each other along the shaft friend. The input channels can be located at a distance from each other or next to each other without a gap, and they can also be located by deploying around the shaft relative to each other.

Additional preferred embodiments of a supercharged internal combustion engine will be disclosed upon consideration of the dependent claims of the utility model.

Preferred are embodiments of a supercharged internal combustion engine in which the supply ducts are configured such that the following ratio is satisfied: 0.9 p p _{1, in} / p _{2, in 1} 1.1, with p _{1, in} denoting the pressure immediately before the first rotor, ap _{2, in} denotes the pressure immediately before the second rotor.

Preferred, in particular, are embodiments of a supercharged internal combustion engine in which the supply ducts are configured such that the following ratio is satisfied: 0.95 p p _{1, in} / p _{2, in} 1 1.05, with p _{1, in} denotes the pressure immediately before the first rotor, and p _{2, in} denotes the pressure immediately before the second rotor.

Preferred, in particular, are embodiments of a supercharged internal combustion engine in which the supply ducts are configured such that the following relationship is satisfied: 0.975≤p _{1, in} / p _{2, in} ≤ 1,025, with p _{1, in} denoting the pressure immediately before the first rotor, and p _{1, in} denotes the pressure immediately before the second rotor.

The more similar are the pressure ratios p _{1, in} / p _{i, out} and the charge air flows in the individual rotors, i.e., the closer they are to each other, the more favorable it is for the compressor operating characteristics, and therefore for the turbocharger and finally, for a supercharged internal combustion engine in general.

Preferred are embodiments of a supercharged internal combustion engine in which the housing comprises a common exhaust channel for said at least two rotors.

The common outlet channel ensures equal pressure p _{i, out} after the individual rotors, as a result of which the implementation of equal in magnitude or close pressure ratios p _in / p _{out is} greatly simplified.

Preferred are embodiments of a supercharged internal combustion engine in which said at least two rotor feed channels have a curved profile starting from their input channels, each feed channel comprising a feed channel in the form of a portion at the end closest to the rotor that forms angle α <45 °.

Preferred, in particular, are embodiments of a supercharged internal combustion engine in which the angle a corresponds to the following: α <20 °.

Preferred, in particular, are embodiments of a supercharged internal combustion engine in which the angle a corresponds to the following: α <10 °.

The greater the axial component of the charge air flow rate on the approach to the rotor, the more favorable this is for the compressor efficiency. In this regard, it is advisable that the feed channel forms at the end of the feed channel closest to the rotor as acute angle α as possible, i.e., so that the angle is as small as possible.

Preferred are embodiments of a supercharged internal combustion engine in which the compressor is a radial compressor. The radial arrangement of the compressor provides a compact design if two or more rotors are provided in the compressor, in particular if two rotors are provided. In addition, the radial type design simplifies, in other words, makes it possible to cool the charge air during compression, in particular in such a way that the charge air to be compressed can be cooled over a large area due to the large details of the rotor protruding radially outwardly, which not so easy to use with an axial compressor, which provides only a small installation space for cooling the charge air in the rotor area. However, an internal combustion engine according to the present utility model may in principle be equipped with an axial type compressor.

In this context, embodiments of a supercharged internal combustion engine are preferred in which each rotor comprises a front side on which rotor blades are located, the front side extending perpendicular to the shaft and intersecting said shaft.

Rotors of this type are used in radial compressors and contribute to compact design, in particular if two rotors are provided, as in the following embodiment.

Preferred, in particular, are embodiments of a supercharged internal combustion engine in which two rotors are mounted on the shaft so that their rear sides are opposed to each other. This design simplifies the use of a common exhaust channel for two rotors.

Preferred in this case are embodiments of a supercharged internal combustion engine in which the rear sides of the two rotors are opposed to each other without a gap. This design also simplifies the use of a common exhaust channel, and additionally provides a compact compressor design.

Preferred are embodiments of a supercharged internal combustion engine in which the compressor is equipped with a liquid cooling system, and for this purpose, the housing comprises at least one built-in refrigerant channel.

A liquid-cooled compressor also has several advantages. Due to liquid cooling, the temperature of the charge air at the outlet of the compressor can be reduced, and the formation of deposits in the compressor is prevented, in particular due to contamination of the ventilation stream. Tests show that the temperature at the compressor outlet significantly affects both the formation of deposits and the volume of deposits. The formation of deposits can be prevented as long as possible until the outlet temperature exceeds a critical value.

Deposits change the geometry of the compressor, in particular, the cross-sectional area of the flows, and thus reduce the efficiency of the compressor. A decrease in compressor efficiency, in turn, leads to an increase in the outlet temperature, which contributes to even more pronounced fouling, i.e. contributes to the formation of deposits.

The cooling of the charge air during compression not only reduces the temperature of the compressed air at the outlet, but also, in particular, increases the efficiency of the compressor. If, for simplicity, we consider the compression process as a stepwise increase in pressure, cooling the air during compression leads to the fact that after each increment of pressure a lower temperature of compressed air will be obtained than in the case of compression without cooling. A lower temperature of the compressed air leads to an increase in the density of compressed air intended for the combustion process in the engine, and thus to a reduction in the volumetric flow rate, which in turn means that a stepwise increase in pressure requires lower compressor power, which causes an increase in efficiency while the same pressure is achieved when the compressor performs less work.

Preferred are embodiments of a supercharged internal combustion engine in which the liquid cooling system is a low temperature circuit.

Due to the low-temperature cooling, the temperature of the charge air can be reduced even at relatively low engine loads and relatively low compression temperatures, for example, 110 ° C. By arranging low temperature cooling, the refrigerant supplied to the duct integrated in the housing can be cooled to temperatures of, for example, 30 ° C, 40 ° C or 50 ° C, so that charging air temperatures of 60 ° C or 70 ° C can be obtained. . Temperatures in this range are also suitable for increasing compressor efficiency.

In this context, embodiments of a supercharged internal combustion engine in which at least one integrated cooling channel is located on the outer circumference of the rotor that is closest to the turbine are preferred. The rotor closest to the turbine is the rotor that carries a higher heat load from the hot exhaust gases, which is why it is preferable to partially cool the compressor, namely this rotor.

Preferred are embodiments of a supercharged internal combustion engine in which the compressor comprises two rotors. Comments have already been made above regarding the use of exactly two rotors.

Preferred are embodiments of a supercharged internal combustion engine in which the shaft is mounted on a support at its end closest to the compressor, on the compressor side, which is facing away from the turbine.

The arrangement and construction of the supply channels and input channels corresponding to the present utility model provide an opportunity to the free compressor end of the shaft. According to the previous embodiment, this is preferably used to provide additional shaft support, thereby improving the vibrational characteristics of the turbocharger as an oscillating system, and therefore, in particular, improving the noise characteristics.

Preferred in this case are embodiments of a supercharged internal combustion engine in which said support is an antifriction bearing or a sliding bearing.

Preferred are embodiments of a supercharged internal combustion engine in which after said at least one compressor a charge air cooler is provided, by means of which compressed charge air is cooled before it enters at least one cylinder. The cooler lowers the temperature and thereby increases the density of charge air, so that the cooler also helps to improve the charge quality in the cylinders, i.e. contributes to an increase in air mass. In fact, compression occurs due to cooling.

If the compressor is cooled by a liquid cooling system, then preferred are embodiments of a supercharged internal combustion engine in which the charge air cooler is located in the low-temperature circuit after the housing.

Preferred are embodiments of a supercharged internal combustion engine in which the compressor housing and / or turbine housing are of a modular design, i.e. the housing consists of several parts, and are composed of at least two sections.

However, embodiments of a supercharged internal combustion engine may also be preferred, in which the compressor housing and / or turbine housing are made in one piece. Then, connecting elements such as, for example, screws and the like are excluded, and assembly of housing parts is also excluded.

Brief Description of the Drawings

Next, a utility model will be described in more detail with an example embodiment, according to FIG. 1 and 2, of which:

FIG. 1 is a cross-sectional view of a turbocharger of a supercharged internal combustion engine according to the prior art, and

FIG. 2 is a cross-sectional view of a turbocharger according to a first embodiment of a supercharged internal combustion engine.

Utility Model Implementation

FIG. 1 is a cross-sectional view of a conventional turbocharger 21 driven by an exhaust gas that is part of a supercharged internal combustion engine and which has already been discussed in relation to the prior art.

FIG. 2 is a cross-sectional view of a turbocharger 1 that is driven by exhaust gas and which corresponds to a first embodiment of a supercharged internal combustion engine. Next, only differences of this embodiment from the embodiment shown in FIG. 1, and for this reason, references in other cases will be made to FIG. 1 and the corresponding description. Components similar to each other, that is, components that perform similar functions, are denoted by the same numbers.

In contrast to the turbocharger 21 shown in FIG. 1, the rotors 2b ₁ , 2b _{2 of the} turbocharger 1 shown in FIG. 2 have similar supply channels 4b ₁ , 4b ₂ , i.e. supply channels 4b ₁ , 4b ₂ that direct charge air to the rotors 2b ₁ , 2b _{2 of} compressor 2, starting from the input channel 5b ₁ , 5b ₂ and up to the rotor 2b ₁ , 2b ₂ at the end of the inlet channel 4b ₁ , 4b ₂ closest to the rotor in the same way, due to the same arrangement of the channels and comparable changes in direction. This provides similar flow ratios in the supply channels 4b ₁ , 4b ₂ and, therefore, similar inlet pressures p _{i, in} immediately in front of the rotors 2b ₁ , 2b ₂ .

The input channels 5b ₁ , 5b _{2 of the} two supply channels 4b ₁ , 4b ₂ extend radially with respect to the shaft 8, and are located next to each other along the shaft 8. In the embodiment shown in FIG. 2, the input channels 5b ₁ , 5b ₂ are located next to each other without a gap.

At the rotor end, each inlet channel 4b ₁ , 4b ₂ contains in the form of a section a feed channel 6b ₁ , 6b ₂ , which is located at an acute angle α <5 ° with respect to the shaft 8. Axial flow incident on the rotors 2b ₁ , 2b ₂ , significantly increases the efficiency of compressor 2.

The arrangement of both input channels 5B1, 5b2 radially with respect to the shaft 8 allows the shaft 8 to be installed with its end 10 closest to the compressor in the support 9 from the side of the compressor 2, which is facing from the turbine 3, as a result of which the dynamic characteristics as well as the vibration characteristics of the shaft 8 are improved and turbocharger 1. In the embodiment shown in FIG. 2, the support 9 is an antifriction bearing 9a.

Item Numbers

1 exhaust gas turbocharger

2 compressor

2a compressor housing

2a ′ spiral housing; compressor output

2b ₁ first compressor rotor

2b ₂ second compressor rotor

3 turbine

3a turbine housing

3a ′ spiral housing; turbine inlet

3b turbine rotor

4b ₁ first inlet channel; inlet channel of the first compressor rotor

4b ₂ second supply channel; inlet channel of the second compressor rotor

5b ₁ first input channel; the input channel of the first compressor rotor 5b _{2 the} second input channel; the input channel of the second rotor of the compressor 6b _{1 the} first supply channel; the feed channel of the first compressor rotor 6b _{2 a} second feed channel; feed channel of the second compressor rotor

7 common exhaust channel

8 shaft

9 prop

9a anti-friction bearing

10 shaft end closest to the compressor

State of the art

21 exhaust gas turbocharger

22 compressor

22a compressor housing

22a ′ spiral housing; compressor output

22b ₁ first compressor rotor

22b ₂ second compressor rotor

23 turbine

23a turbine housing

23a ′ spiral housing; turbine inlet

23b turbine rotor

24b ₁ first inlet channel; inlet channel of the first compressor rotor

24b ₂ second supply channel; inlet channel of the second compressor rotor

25b ₁ first input channel; input channel of the first compressor rotor

25b ₂ second input channel; input channel of the second compressor rotor

26b ₁ first feed channel; feed channel of the first compressor rotor

26b ₂ second supply channel; feed channel of the second compressor rotor

27 common exhaust channel

28 shaft

I number of compressor rotors

p _{1, in} inlet pressure immediately before the first rotor

p _{2, in} inlet pressure immediately in front of the second rotor

p _{i, in} inlet pressure immediately before the i-th rotor

p _{1, out} outlet pressure immediately after the first rotor

p _{2, out} outlet pressure immediately after the second rotor

p _{i, out the} outlet pressure immediately after the i-th rotor.

Claims (16)

1. A supercharged internal combustion engine comprising: - an intake system for supplying an internal combustion engine with charge air, - an exhaust system for exhaust gas and - at least one exhaust gas driven turbocharger (1), in which a compressor (2) located in the intake system and an axial turbine (3) located in the same rotary shaft (8) are installed the exhaust system, the compressor (2) comprising a housing (2a) and at least two rotors (2b ₁ , 2b ₂ ) located in the housing (2a), which are mounted on the shaft (8) and are configured to be separately charged with charge air through the intake system through the feed channel s (4b ₁ , 4b ₂ ), which are separated from each other and made as one piece with the body (2A), characterized in that - the supply channels (4b ₁ , 4b ₂ ) of at least two rotors (2b ₁ , 2b ₂ ) contain in the form of sections input channels (5b ₁ , 5b ₂ ), which are located next to each other along the shaft (8) and extend radially relative to the shaft (8). 2. A supercharged internal combustion engine according to claim 1, characterized in that the supply channels (4b ₁ , 4b ₂ ) are made in such a way that the following is observed: 0.9≤p _{1, in} / p _{2, in} ≤1,1 wherein p _{1, in} denotes the pressure immediately before the first rotor (2b ₁ ), and p _{2, in} denotes the pressure immediately before the second rotor (2b ₂ ). 3. A supercharged internal combustion engine according to claim 1 or 2, characterized in that the supply channels (4b ₁ , 4b ₂ ) are made in such a way that the following is observed: 0.95≤p _{1, in} / p _{2, in} ≤1 , 05, wherein p _{1, in} denotes the pressure immediately before the first rotor (2b ₁ ), and p _{2, in} denotes the pressure immediately before the second rotor (2b ₂ ). 4. A supercharged internal combustion engine according to claim 1 or 2, characterized in that the housing (2a) has a common exhaust channel (7) for at least two rotors (2b ₁ , 2b ₂ ). 5. A supercharged internal combustion engine according to claim 1 or 2, characterized in that at least two supply channels (4b ₁ , 4b ₂ ) of the rotors (2b ₁ , 2b ₂ ) have a curved profile starting from their input channels (5b ₁ , 5b ₂ ), while each supply channel (4b ₁ , 4b ₂ ) contains, in the form of a section at the end closest to the rotor, a supply channel (6b ₁ , 6b ₂ ), which forms an angle α <45 ° with the shaft (8) . 6. The supercharged internal combustion engine according to claim 5, characterized in that the angle α meets the following criterion: α <20 °. 7. The supercharged internal combustion engine according to claim 5, characterized in that the angle α meets the following criterion: α <10 °. 8. The supercharged internal combustion engine according to any one of paragraphs. 1, 2, 6 and 7, characterized in that the compressor (2) is a radial compressor. 9. The supercharged internal combustion engine according to any one of paragraphs. 1, 2, 6 and 7, characterized in that each rotor (2b ₁ , 2b ₂ ) has a rear side on which the rotor blades are located. 10. A supercharged internal combustion engine according to claim 9, characterized in that the two rotors (2b ₁ , 2b ₂ ) are mounted on the shaft (8) so that their rear sides are located opposite each other. 11. The supercharged internal combustion engine according to claim 10, characterized in that the rear sides of the two rotors (2b ₁ , 2b ₂ ) are opposed to each other without a gap. 12. The supercharged internal combustion engine according to any one of paragraphs. 1, 2, 6, 7, 10 and 11, characterized in that the compressor (2) is equipped with a liquid cooling system, for which the housing (2a) has at least one channel made for a single refrigerant. 13. The supercharged internal combustion engine according to claim 12, characterized in that the liquid cooling system is a low-temperature circuit. 14. A supercharged internal combustion engine according to claim 12, characterized in that at least one refrigerant channel formed as a single unit is located on the outer circumference of the rotor (2b ₂ ), which is closest to the turbine (3). 15. The supercharged internal combustion engine according to any one of paragraphs. 1, 2, 6, 7, 10, 11, 13 and 14, characterized in that the compressor (2) contains two rotors (2b ₁ , 2b ₂ ). 16. The supercharged internal combustion engine according to any one of paragraphs. 1, 2, 6, 7, 10, 11, 13 and 14, characterized in that the shaft (8) is installed in the support (9) at its end closest to the compressor (10) on the side of the compressor that is facing away from the turbine (3 )

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