WO2010142287A1

WO2010142287A1 - Compressor impeller

Info

Publication number: WO2010142287A1
Application number: PCT/DE2010/050001
Authority: WO
Inventors: Hannes Benetschik; Sven Eisenbach; Thomas Winter
Original assignee: Man Diesel & Turbo Se
Priority date: 2009-06-08
Filing date: 2010-01-19
Publication date: 2010-12-16
Also published as: DE102009024568A1; JP2012529585A; KR101369601B1; CN102803739A; KR20120036932A; EP2440791A1; CN102803739B

Abstract

The invention relates to a compressor impeller (20) having a plurality of impeller passages (23) for conducting through a fluid to be compressed, wherein each impeller passage has a fluid inlet end (23a) and a fluid outlet end (23b), wherein the respective impeller passages at the fluid inlet end have a first cross-sectional surface (AE) and at the fluid outlet end have a second cross-sectional surface (AA), and wherein a ratio of sizes (GV) between the second cross-sectional surface and the first cross-sectional surface is smaller than 0.7.

Description

compressor impeller

The invention relates generally to a compressor impeller, and more particularly to a compressor impeller for a centrifugal compressor, a centrifugal compressor equipped with such a compressor impeller, and an exhaust gas turbocharger of an internal combustion engine equipped with such a centrifugal compressor.

The lecture script "Kraft- und Arbeitsmaschinen" (University of Applied Sciences)

Schmalkalden, Department of Mechanical Engineering, Version 08/2007) by Prof. Dr.-Ing. Robert Pietzsch deals on pages 46 to 57 with conventional turbocompressors. On page 47 of this script there is described a compressor impeller having a plurality of impeller passages for passing a fluid to be compressed, the impeller passages each having a fluid inlet end and a fluid outlet end, and wherein the respective impeller passages have a first cross-sectional area at the fluid inlet end and a second cross-sectional area at the fluid outlet end.

The energy conversion process in a centrifugal compressor takes place by means of a rotating blading of a compressor impeller, which impels a spin to the pumped medium or fluid to be compressed, such as, for example, atmospheric fresh air. The power of the swirling flowing fluid behaves on the one hand proportional to the peripheral speed of rotating as a solid blading and on the other hand proportional to the concentric rotating component of its speed, which in turn has proportional to the rotational speed of the solid and thus to the speed of the compressor impeller. The throughput promoted with the centrifugal compressor also behaves proportionally to the rotational speed of the rotating blading for flow velocities that can be considered low compared to the speed of sound of the fluid. As the rotational speed of the compressor impeller increases, however, the compressibility of the fluid results in a flow rate limit that depends primarily on the flow areas within the blading. Accordingly, in conventional centrifugal compressors, the pressure, which is characterized as volume-specific energy, behaves disproportionately in relation to the throughput.

Fig. 1 illustrates, showing a total pressure ratio π _to t via

the throughput or volume flow V the typical working range of a conventional centrifugal compressor.

When using the radial compressor in an exhaust gas turbocharger for internal combustion engines, the radial compressor as shown in Figure 1 follows a characteristic (according to the line BC in Figure 1) of the internal combustion engine, but over the entire load range away a minimum distance to a pumping limit PG of the centrifugal compressor ensure. On the other hand, the throughput of the centrifugal compressor is limited by its blocking limit or swallow limit SG. Lines of constant speed of the rotating blading or of the compressor impeller are designated in FIG. 1 with n _k onst.

In conjunction with the one best efficiency or optimal

Efficiency characterizing line WG _op t shows that the disproportionate dependence of the compression ratio or total pressure ratio

Il dead from throughput V in conventional centrifugal compressors to a conflict of goals

with the characteristic BC of the internal combustion engine leads. The to be observed with regard to a stable Betheb of the centrifugal compressor distance from the pumping limit PG has the consequence that both the achievable on the internal combustion engine boost pressure and the thermodynamic efficiency at the operating point significant loss in comparison with the optimal design values of the centrifugal compressor in appearance let go.

Taking into account the material technology limited solid-state rotational speed of the compressor impeller on the outer section or outer periphery of which limit conventional designs therefore the single-stage turbocharging to about 5.5 bar.

In a one-stage charging of an internal combustion engine, such as e.g. a diesel engine, with pressure ratios exceeding a value of about 5.5 bar, operates a conventional centrifugal compressor with efficiencies well below its optimum.

The invention has for its object to provide a compressor impeller for a centrifugal compressor, with the boost pressure or compressor outlet side pressures of over 5.5 bar can be achieved with significantly improved efficiency. The invention is further based on the object to provide a equipped with such a compressor wheel radial compressor and equipped with such a centrifugal compressor exhaust gas turbocharger an internal combustion engine.

The above objects are achieved with a compressor impeller according to claim 1, a centrifugal compressor according to claim 8 and an exhaust gas turbocharger of an internal combustion engine according to claim 11. Further developments of the invention are defined in the respective dependent claims.

According to a first aspect of the invention, a compressor impeller for a radial compressor comprises a plurality of impeller passages for passing a fluid to be compressed, the impeller passages each having a fluid inlet end and a fluid outlet end and wherein the respective impeller passages at the fluid inlet end have a first cross sectional area and at the fluid exit end a second cross sectional area to have. The compressor impeller according to the invention is characterized in that a size ratio of the second cross-sectional area to the first cross-sectional area is smaller than 0.7.

According to the invention, it has been recognized that in conventional compressor impellers the size ratio of the second cross-sectional area to the first cross-sectional area is approximately 0.75, ie in any case greater than 0.7.

The inventive design of Fluideinthttsende and fluid outlet end with an area ratio or size ratio of less than 0.7, the operating characteristics of a equipped with the compressor impeller radial compressor according to the invention are improved so that the best efficiency of such a centrifugal compressor characteristic line in the map almost coincident with or nearly runs parallel to and very close to a line describing the characteristic of an engine operatively coupled to the centrifugal compressor.

Thus, the conflict of objectives between compressor design and turbocharging of the internal combustion engine is largely eliminated. In other words, the disproportionate dependence of the boost pressure on the throughput in only slightly pronounced form is evident, so that the radial compressor can follow the characteristic of the internal combustion engine following its optimal design. The solution according to the invention can lead to compressor maps at the expense of a moderate theoretical maximum pressure, which have reduced throughputs, in particular in the partial load range, but with a significant increase in the charge pressure that can actually be provided and

Compressor efficiency is achieved. This is achieved in particular by the inventive design of the ratio of the cross-sectional area of the impeller passages at the fluid end end to the cross-sectional area of the impeller passages at the fluid end end.

A centrifugal compressor equipped with a compressor impeller according to the invention has a broader stable working range which can be seen in the characteristic field, resulting in reduced throughputs under partial load conditions, so that the radial compressor and the characteristic of an internal combustion engine operatively coupled thereto are matched to one another in the best possible way.

According to one embodiment of the compressor impeller according to the invention, the size ratio of the second cross-sectional area to the first cross-sectional area is smaller than 0.65.

In such an embodiment of the compressor impeller according to the invention are essentially without loss of efficiency in the operating point boost pressures to the internal combustion engine of up to 6 bar realized.

According to one embodiment of the compressor impeller according to the invention, a size of the first cross-sectional area is at least 1.54 times a size of the second cross-sectional area. In such an embodiment of the compressor impeller according to the invention are essentially without loss of efficiency in the operating point boost pressures to the internal combustion engine of up to 6 bar realized.

According to a further embodiment of the compressor impeller according to the invention, the size ratio of the second cross-sectional area to the first cross-sectional area is smaller than 0.6.

With such an embodiment of the compressor impeller according to the invention, boost pressures to the internal combustion engine of up to 7 bar can be realized substantially without sacrificing the efficiency at the operating point.

According to one embodiment of the compressor impeller according to the invention, a size of the first cross-sectional area is at least 1.67 times a size of the second cross-sectional area.

According to yet another embodiment of the compressor impeller according to the invention, the compressor impeller further comprises an impeller hub having an outer circumference and a plurality of impeller blades distributed along the outer circumference of the impeller hub on the impeller hub and each having two lateral blade surfaces and a radially outer one disposed between the blade surfaces Have edge. According to this embodiment of the compressor impeller according to the invention, the outer edges of the impeller together define an outer periphery of the impeller blades, the respective impeller passages being formed between respective adjacent impeller blades. The impeller passages are in each case of the outer circumference of the impeller hub, opposite blade surfaces of the respective adjacent impeller seh on one and the outer periphery of the impeller seh limited to a. According to this embodiment of the invention, the fluid inlet end of the respective impeller passages is arranged radially inwardly and the fluid end of the respective impeller passages is arranged radially outboard.

According to another embodiment of the compressor impeller according to the invention, the impeller passages between the fluid inlet end and the fluid outlet end each have a partition dividing the impeller passage after the fluid inlet end into two sub-passages, the partition wall extending along the impeller passage from a separation location a predetermined distance from the fluid inlet end , extends to the fluid exit end so that the impeller passage has a single fluid inlet and two fluid outlets juxtaposed in a circumferential direction of the compressor impeller.

In particular, such a partition has the advantages that on the one hand it does not reduce the cross-sectional area of the respective impeller passage at the fluid inlet end and on the other hand causes a better transfer of the mechanical work or kinetic energy performed by the compressor impeller to the fluid to be compressed.

According to one embodiment of the compressor impeller according to the invention, the two partial passages of each impeller passage at the fluid outlet end each have an outlet cross-sectional area, wherein a sum of a respective Size of the respective outlet cross-sectional areas of the two partial passages is equal to a size of the second cross-sectional area.

This embodiment of the invention provides additional flexibility or design freedom for an optimal design of the compressor impeller, since the two outlet cross-sectional areas of the two partial passages of each impeller passage can be made the same size or different sizes as needed.

Preferably, the partitions are formed by respective auxiliary blades, which correspond in their shape and their radial extent to the impeller blades and which are formed fluideinthendseitig by a predetermined distance corresponding dimension shorter than the impeller blades.

This embodiment of the compressor impeller according to the invention supports in a particularly advantageous manner, the transmission of the mechanical work done by the compressor impeller or the kinetic energy to the fluid to be compressed.

According to a second aspect of the invention, there is provided a radial compressor for a turbocharger, the radial compressor comprising a compressor housing having a fluid inlet for receiving a main flow of fluid to be compressed in the centrifugal compressor and a feedback passage, and a compressor wheel according to one or more or all of those previously described Embodiment (s) of the invention, wherein the compressor impeller is rotatably mounted in the compressor housing disposed in a flow direction of the main flow after the fluid inlet, and wherein the feedback passage from a fluid inlet located at the first inner peripheral portion of the compressor housing to a radially surrounding the compressor impeller second inner peripheral portion the compressor housing extends so that along the feedback passage can form a side stream of fluid to be compressed.

Due to the inventive design of fluid inlet end and fluid outlet end of the compressor impeller of the invention

Radial compressor, the operating characteristics of the centrifugal compressor are improved so that the best efficiency of the centrifugal compressor characteristic line in the map almost coincident with or nearly parallel to and very close to a characteristic of an operatively coupled to the centrifugal internal combustion engine line runs.

Thus, the target conflict between radial compressor design and turbocharging the engine is largely eliminated. In other words, the disproportionate dependence of the boost pressure on the throughput in only slightly pronounced form is evident, so that the radial compressor can follow the characteristic of the internal combustion engine following its optimal design. Furthermore, can be dispensed with by the inventively provided solution if necessary on Abblaseklappen in the centrifugal compressor.

The solution according to the invention can lead to radial compressor characteristic diagrams, which have reduced throughputs, in particular in the partial load range, at the expense of a moderate theoretical maximum pressure, whereby, however, a significant increase in the actual charge pressure and the efficiency of the radial compressor is achieved. This is achieved in particular by the inventive design of the ratio of the cross-sectional area of the impeller passages at the fluid outlet end to the cross-sectional area of the impeller passages at the fluid inlet end of the compressor impeller.

The radial compressor according to the invention has a broader stable working range which can be seen in the characteristic diagram, which has the result, together with reduced throughputs under partial load conditions, that the Radial compressor and the characteristics of a so operatively coupled internal combustion engine are almost perfectly matched.

Due to the fact that the compressor housing of the radial compressor according to the invention has the feedback passage which permits a compressor-internal recirculation, the greatest possible proportionality between the obtained pressure ratio or total pressure ratio and the throughput or volume flow along the surge line is supported.

In other words, the configuration of the recirculation space or the feedback passage at the surge line permits a feedback and reunification of a fluid flow torn off from the outer circumference of the impeller blades (or recycling these as a secondary flow against the flow direction of the main flow) to the main flow upstream of the compressor impeller, which ensures that a stabilization of the flow conditions in the compressor impeller can be established.

On the other hand, the feedback passage can be designed such that a secondary portion or secondary flow of the conveying medium or fluid to be compressed downstream of the first cross-sectional area relevant for the throughput, in particular at high compressor speeds in the flow direction of the main flow via the feedback passage, reunites with the main flow of the fluid.

That is, with the design and arrangement of the feedback passage, the support of a stable compressor operation can be concentrated on the surge limit or on the stop limit, but any variations or gradations are possible between these two extreme embodiments. According to an embodiment of the radial compressor according to the invention, in the feedback passage, guide vanes are preferably arranged for influencing a flow direction and / or a flow rate of the secondary flow.

Preferably, the feedback passage has a first end located at the fluid inlet and a second end located near the fluid inlet end of the impeller passages.

According to one embodiment of the radial compressor according to the invention this is designed as a single-stage centrifugal compressor.

According to a third aspect of the invention, there is provided an exhaust gas turbocharger of an internal combustion engine having an exhaust gas turbine and a centrifugal compressor according to one or more or all of the above-described embodiment (s) of the invention.

A vehicle equipped with such an exhaust gas turbocharger internal combustion engine (internal combustion engine) of a vehicle has in particular a higher power output and lower fuel consumption.

In the following the invention will be described in more detail by means of preferred embodiments and with reference to the attached figures.

1 shows a characteristic diagram of the working range of a conventional radial compressor.

2 shows a schematic view of a radial compressor according to an embodiment of the invention.

3 shows a schematic view of a compressor impeller of the radial compressor of Figure 2. 4 shows a characteristic diagram of the working range of the radial compressor of FIG.

FIG. 5 shows a view in which the characteristic diagrams of FIGS. 1 and 4 are shown superimposed for comparison purposes.

In the following, an exhaust gas turbocharger of an internal combustion engine (not completely shown) according to embodiments of the invention will be described with reference to the accompanying figures 2-5.

The exhaust gas turbocharger according to the invention has an exhaust gas turbine (not shown), which is connected on the input side to an exhaust system of a motor vehicle (not shown) of a motor vehicle (not shown) and a single-stage radial compressor 1 (shown in FIG. 2 and FIG. 3). on, which is drehanthebsverbunden connected via a drive shaft, not shown, with the exhaust turbine.

The centrifugal compressor 1 comprises a compressor housing 10 having a fluid inlet 11 for receiving a main flow H of optionally filtered atmospheric fresh air to be compressed in the centrifugal compressor 1, a fluid outlet (not shown) for discharging the compressed fresh air, the fluid outlet having an air inlet of the internal combustion engine fluidly connected, and having a feedback passage 12.

The centrifugal compressor 1 further has a compressor impeller 20, which is rotatably mounted in the compressor housing 10 in a direction of flow of the main stream H after the fluid inlet 11, indicated by the arrowhead of the main stream H symbolizing line.

The feedback passage 12 extends from a first inner peripheral portion 13 of the compressor housing 10 located at the fluid inlet 11 to a second radially surrounding the compressor wheel 20 Inner peripheral portion 14 of the compressor housing 10, so that along the feedback passage 12, a side stream N to be compressed fresh air depending on the operating conditions opposite to or corresponding to the flow direction of the main flow H.

The feedback passage 12 is formed by an annular recess 15 in the inner circumference of the compressor housing 10 and a ring member 16 inserted into the fluid inlet 11 so that the feedback passage 12 has a first end 12a located at the fluid inlet 11 and a second end 12b is near respective fluid inlet ends 23a of impeller passages 23 of the compressor impeller 20.

For influencing the flow direction and / or the flow rate of the secondary flow N, fixed or adjustable guide vanes 17 are arranged in the feedback passage 12.

A with regard to their operation of the feedback passage 12 corresponding feedback passage is described for example in DE 33 22 295 C3.

The compressor impeller 20 has a rotatably mounted on the drive shaft impeller hub 21 with an outer circumference 21 a and a plurality of impeller seh on a 22 along the outer periphery 21 a of the impeller hub 21 in the circumferential direction evenly distributed on the impeller hub 21 are arranged and each two lateral blade surfaces 22a and 22b and a radially outer edge 22c extending between the two blade surfaces 22a, 22b.

The outer edges 22c of the respective impeller blades 22 together define an outer periphery (considered as a rotational body) of the Impeller sch on a 22, wherein between each adjacent impeller seh on a 22, 22 each have an impeller passage 23 for passing the fresh air to be compressed (fluid) is formed.

The thus formed impeller passages 22 each have a radially inner (near the drive shaft disposed) Fluideinthttsende 23a and a radially outboard (radially more than the fluid inlet end 23a of the drive shaft spaced) Fluidausthttsende 23b. The outer periphery of the impeller sch on a 22 is enclosed with a small gap therebetween from the radially outside of the second inner peripheral portion 14 and the ring member 16 of the compressor housing 10.

The impeller passages 23 are each of the outer circumference 21 a of the impeller hub 21, opposite blade surfaces 22 a, 22 b of the respective adjacent impeller sch on a 22, 22 and the outer periphery of

Impeller sch limited to a 22 and the second inner peripheral portion 14 and the ring member 16 of the compressor housing 10.

The impeller passages 23 have, between their respective fluid inlet end 23a and their respective fluid end 23b, a partition in the form of an additional blade 24, which in its radial extent coincides with the impeller 52, but is formed shorter than the impeller blades 22 by a certain amount is.

In other words, each impeller passage 23 is subdivided into two subpassages 23c, 23d downstream of its fluid inlet end 23a, whereby the additional vane 24 acting as a partition wall extends along the impeller passage 23 from a separation point T (only indicated in FIG. 3) which is a predetermined distance from the fluid inlet end 23a has extended to the fluid end 23b, so that the impeller passage 23 a single fluid inlet and two in one Circumferential direction of the compressor impeller 20 has adjacent fluid outputs.

Each impeller passage 23 has at its fluid inlet end 23a a first cross-sectional area or inlet cross-sectional area AE.

The two partial passages 23c, 23d of each impeller passage 23 each have an outlet cross-sectional area AAi or AA ₂ at the fluid outlet end 23b of the respective impeller passage 23. The according to this embodiment, the same size two outlet cross-sectional areas AA-i, AA _{2 of} the two

Partial passages 23c, 23d in their sum form a second cross-sectional area or total exit cross-sectional area AA of the respective impeller passage 23 (AA = AAi + AA ₂ ). According to further embodiments of the invention, the two outlet cross-sectional areas AA-i, AA _{2 can} also be of different sizes.

According to one embodiment of the invention, the size ratio GV of the second cross-sectional area AA to the first cross-sectional area AE is smaller than 0.7, the following rule being true:

GV = AA / AE.

With a size ratio GV of less than 0.7, supercharging pressures on the internal combustion engine of greater than 5.5 bar can be achieved without sacrificing the efficiency at the operating point.

According to a further embodiment of the invention, the size ratio GV of the second cross-sectional area AA to the first cross-sectional area AE is smaller than 0.65. With a size ratio GV of less than 0.65, supercharging pressures on the internal combustion engine of up to 6 bar can be achieved without sacrificing efficiency at the operating point.

According to yet another embodiment of the invention, the size ratio GV of the second cross-sectional area AA to the first cross-sectional area AE is less than 0.6.

With a size ratio GV of less than 0.6 bar can be realized without sacrificing efficiency at the operating point boost pressures on the engine of up to 7 bar.

4 shows a characteristic diagram of the working range of the radial compressor 1 according to the invention. FIG. 5 shows a view in which the characteristic diagrams of FIGS. 1 and 4 are shown superimposed for comparison purposes.

As can be seen from FIG. 4 and FIG. 5 (in conjunction with FIG. 1), the inventive design of the size ratio GV of fluid end 23a and fluid end 23b of the compressor impeller 20 of the radial compressor 1 according to the invention improves the operating characteristics of the radial compressor 1 in this way in that a line WG ' _op t characterizing the best efficiency of the centrifugal compressor 1 runs nearly coincident with or nearly parallel to and very close to a line BC describing the characteristic of the internal combustion engine operatively coupled to the centrifugal compressor 1 in the characteristic map.

Thus, the well-known from the prior art conflict between the radial compressor design and turbocharging the engine is largely eliminated. In other words, the disproportionate dependence of the boost pressure or total pressure ratio IT _to t of the throughput occurs or volumetric flow V in only slightly pronounced form so that the radial compressor 1 can follow the characteristic BC of the internal combustion engine following its optimum design.

The radial compressor 1 according to the invention has, assuming a moderate theoretical maximum pressure (top end of the lines WG _op t, WG'o _p t) on a map, which reduced especially in the partial load range

Throughputs or volume flows V shows, but with a significant increase

Δπ dead of the actual supplyable boost pressure and the efficiency of the radial compressor 1 is achieved.

The radial compressor 1 according to the invention has an apparent in the map wider stable working area, resulting in interaction with

under partial load conditions reduced throughputs or volume flows V has the consequence that the radial compressor 1 and the characteristic BC of the thus operatively coupled internal combustion engine are matched almost optimally to each other.

The fact that the compressor housing 10 of the radial compressor 1 according to the invention has the feedback passage 12, which allows a compressor internal recirculation, a maximum proportionality between the achieved pressure ratio or

Total pressure ratio π _to t and the flow rate or volume flow V along the

Pump limit PG 'supported.

The design of the vanes 17 and the dimensioning / arrangement of the feedback passage 12 allow feedback at the pumping limit PG ' and reunification of a fresh air flow torn off from the outer circumference of the impeller onto a 22 and the additional blades 24 (or return of these as bypass flow N against the flow direction of the main flow H) to or with the main flow H upstream of the compressor impeller 20, which ensures that can set a stabilization of the flow conditions in the compressor impeller 20.

On the other hand, the design of the vanes 17 and the dimensioning / arrangement of the feedback passage 12 make it possible for the feedback passage 12 to be arranged so as to be particularly at high

Compressor speeds in the flow direction of the main flow H via the feedback passage 12 flowing side stream N of the fresh air to be compressed downstream of the flow rate or volume flow

V relevant first cross-sectional area AE with the main flow H of the fresh air to be compressed reunited.

That is, with the configuration of the feedback passage 12, the assistance of a stable compressor operation can be concentrated on the surge limit PG 'or on the stop limit SG', however, any variations or gradations are possible between these two extreme embodiments.

LIST OF REFERENCE NUMBERS

1 radial compressor

10 compressor housing 11 fluid inlet

12 feedback passage 12a first end

12b second end

13 first inner peripheral portion 14 second inner peripheral portion

15 recess

16 ring element

17 vane (s) 20 Compressor impeller 21 impeller hub

21 a outer circumference

22 impeller blade (s) 22a lateral blade surface 22b lateral blade surface 22c radially outer edge

23 Impeller passage (s) 23a Fluid inlet end 23b Fluid exit end 23c Partial passage (s) 23d Partial passage (s)

24 additional bucket (s)

T separation point

AE first cross-sectional area (inlet cross-sectional area) AA second cross-sectional area (total outlet cross-sectional area)

AAi outlet cross-sectional area

AA ₂ outlet cross-sectional area H mainstream

N sidestream

II dead total pressure ratio

Δπ _dead increase

V flow rate

PG surge limit

SG blocking limit

WGopt optimum efficiency

BC characteristic of the internal combustion engine ⁿ const constant speed

PG ¹ surge limit

SG ¹ blocking limit

WG'opt optimum efficiency

BC characteristic of the internal combustion engine ⁿ const constant speed

Claims

claims

A compressor impeller (20) for a radial compressor 1, comprising a plurality of impeller passages (23) for passing a fluid to be compressed, the impeller passages (23) each having a fluid inlet end (23a) and a

Have fluid outlet end (23b), and wherein the respective impeller passages (23) at the fluid inlet end (23a) has a first cross-sectional area (AE) and the fluid outlet end (23b) has a second cross-sectional area (AA), characterized in that a size ratio (GV) of second

Cross-sectional area (AA) to first cross-sectional area (AE) is less than 0.7.

Second compressor impeller (20) according to claim 1, wherein the size ratio (GV) of the second cross-sectional area (AA) to first cross-sectional area (AE) is less than 0.65.

3. compressor impeller (20) according to claim 1 or 2, wherein the size ratio (GV) of the second cross-sectional area (AA) to first cross-sectional area (AE) is less than 0.6.

4. compressor impeller (20) according to one of claims 1 to 3, further comprising: an impeller hub (21) having an outer circumference (21 a), a plurality of impeller blades (22) along the outer periphery (21 a) of the impeller hub (21 ) are arranged distributed on the impeller hub (21) and the respective two lateral blade surfaces (22a, 22b) and one between the

Vane surfaces (22a, 22b) disposed radially outer edge (22c), wherein the outer edges (22c) of the impeller blades (22) together define an outer periphery of the impeller blades (22), wherein the respective impeller passages (23) between each adjacent impeller blades (22, 22) are formed, wherein the impeller passages (23) respectively from the outer periphery (21 a) of the impeller hub (21), opposite blade surfaces (22 a, 22 b) of the respectively adjacent Impeller blades (22, 22) and the outer periphery of the impeller blades (22) are limited, and wherein the Fluideinthttsende (23a) of the respective impeller passages (23) is arranged radially inwardly and the Fluidusthtende (23b) of the respective impeller passages (23) is arranged radially outboard ,

5. compressor impeller (20) according to one of claims 1 to 4, wherein the impeller passages (23) between the fluid inlet end (23 a) and the fluid outlet end (23 b) each having a partition wall, the impeller passage (23) after the fluid inlet end (23 a) in two sub-passages (23c, 23d) are partitioned, wherein the partition wall extends along the impeller passage (23) from a separation point (T) having a predetermined distance to the fluid inlet end (23a) to the fluid exit end (23b), so that the impeller passage (23) has a single fluid inlet and two adjacent in a circumferential direction of the compressor impeller (20) fluid outputs.

6. compressor impeller (20) according to claim 5, wherein the two partial passages (23 c, 23 d) each impeller passage (23) at the fluid outlet end (23 b) each have an outlet cross-sectional area (AA-i, AA ₂ ), and wherein a sum of a respective size of respective outlet cross-sectional areas (AA-i, AA ₂ ) of the two partial passages (23c, 23d) is equal to a size of the second cross-sectional area (AA).

7. compressor impeller (20) according to claim 5 or 6, wherein the partitions of respective auxiliary blades (24) are formed, which in its radial extent to the impeller on a (22) correspond and which Fluid inlet end side by a predetermined distance corresponding dimension shorter than the impeller blades (22) are formed.

A centrifugal compressor (1) for a turbocharger, comprising: a compressor housing (10) having a fluid inlet (11) for receiving a main flow (H) of fluid to be compressed in the radial compressor (1) and a feedback passage (12) A compressor impeller (20) according to any one of claims 1 to 7, wherein the compressor impeller (20) is rotatably mounted in the compressor housing (10) in a flow direction of the main flow (H) downstream of the fluid inlet (11), the feedback passage (12) from a first inner peripheral portion (13) of the compressor housing (10) located at the fluid inlet (11) to a second inner peripheral portion (14) of the compressor housing (10) radially enclosing the compressor impeller (20) so that a bypass flow is provided along the feedback passage (12) (N) can flow of fluid to be compressed.

9. Radial compressor (1) according to claim 8, wherein in the feedback passage (12) vanes (17) are arranged for influencing a flow direction and / or a flow rate of the secondary flow (N).

A centrifugal compressor (1) according to claim 8 or 9, wherein the feedback passage (12) has a first end (12a) located at the fluid inlet (11) and a second end (12b) located near the fluid entrance end (23a) of the impeller passages (23) is located.

11. Exhaust gas turbocharger of an internal combustion engine, with an exhaust gas turbine and a centrifugal compressor (1) according to one of claims 8 to 10.