JP2012529585A - Compressor impeller - Google Patents

Abstract

  A compressor impeller (20) having a plurality of impeller passages (23) for allowing a fluid to be compressed to flow, the impeller passages each having one fluid inflow end (23a) and one fluid outflow end (23b). And each impeller passage has a first cross section (AE) at a fluid inflow end, a second cross section (AA) at a fluid outflow end, and a second cross section with respect to the first cross section. The size ratio (GV) is less than 0.7.

Description

  The present invention generally relates to a compressor impeller for a compressor impeller, in particular a centrifugal compressor, a centrifugal compressor equipped with such a compressor impeller, and an internal combustion engine equipped with such a centrifugal compressor Related to the exhaust gas turbocharger.

  Non-Patent Document 1 deals with conventional turbo compressors on pages 46 to 57. Page 47 of this manuscript describes a compressor impeller having a plurality of impeller passages for allowing the fluid to be compressed to flow therethrough, each impeller passage having a fluid inflow end and a fluid outflow end. Each impeller passage has a first cross section at the fluid inflow end and a second cross section at the fluid outflow end.

  The energy conversion process in a centrifugal compressor takes place by a rotating cascade of compressor impellers that apply a swirling motion to a transfer medium or fluid to be compressed, such as atmospheric ambient air. The performance of the fluid flowing under the swirl movement is then proportional to the circumferential speed of the cascade rotating as a solid and on the other hand proportional to the concentric rotational component of the solid velocity. In relation, the component is proportional to the rotational speed of the solid and thus to the rotational speed of the compressor impeller.

  The flow rate conveyed by the centrifugal compressor is similarly proportional to the number of revolutions of the rotating blade row at a flow rate that can be considered low compared to the acoustic velocity of the fluid. However, as the rotational speed of the compressor impeller increases, the compression rate of the fluid results in a limit value that first depends on the flow cross section inside the blade cascade. Accordingly, in conventional centrifugal compressors, the pressure characterized as volumetric specific energy is clearly overly dependent, depending on the flow rate.

FIG. 1 shows the typical working area of a conventional centrifugal compressor, representing the total pressure ratio Π _tot for flow rate or volume flow rate V. In addition, the symbol V indicating “volume flow rate” described in the present specification is the symbol shown on the horizontal axis in the graphs shown in FIGS. 1, 4, and 5 (“·” (dot) immediately above “V”). It should be noted that it will be used in the following as an alternative to

When using a centrifugal compressor in an exhaust gas turbocharger for an internal combustion engine, the centrifugal compressor follows the characteristic curve of the internal combustion engine (according to line BC in FIG. 1), as shown in FIG. However, a minimum spacing for the centrifugal compressor pump limit PG should be guaranteed over the full load range. On the other hand, the flow rate of the centrifugal compressor is limited by its shut-off limit or absorption limit or blockage limit SG. A constant rotational speed line of a rotating blade row or compressor impeller is denoted by n _konst in FIG.

If the compression pressure ratio or the total pressure ratio Π _tot is excessively dependent on the flow rate V in the overview with the line WG _opt indicating the highest efficiency or the optimum efficiency, the characteristic curve BC of the internal combustion engine is Has been shown to lead to conflicts of goals.

  The distance from the pump limit PG, which should be maintained in consideration of the stable drive of the centrifugal compressor, is particularly high under the load, the supply pressure achievable with an internal combustion engine, the thermodynamic efficiency at the point of action, The result is a clear loss compared to the optimum design value of the centrifugal compressor.

  Therefore, in view of the solid rotational speed of the compressor impeller, which is limited in terms of material technology, at the outer cross-section or outer periphery of the compressor impeller, the conventional design provides a single stage turbocharge of approximately 5,5. Limited to bar.

  In a single-stage supercharging of an internal combustion engine, such as a diesel engine, having a pressure ratio exceeding a value of approximately 5,5 bar, a normal centrifugal compressor operates at an efficiency clearly below its highest level.

German Patent No. 3322295

Dr. Lecture manuscript by Prof. Roberto Pecchu “Motor and Actuator” (Schmalkarden University, Department of Mechanical Engineering, August 2007 edition)

  The object of the present invention is to provide a compressor impeller for a centrifugal compressor, which can achieve a supply pressure of more than 5,5 bar or a pressure on the outlet side of the compressor with a clearly improved efficiency. is there. It is a further object of the present invention to provide 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 above problem is solved by the compressor impeller according to claim 1 and the centrifugal compressor according to claim 8 or the exhaust gas turbocharger of the internal combustion engine according to claim 11. Further aspects of the invention are defined in the respective dependent claims.

  According to the first aspect of the present invention, a compressor impeller for a centrifugal compressor includes a plurality of impeller passages for allowing a fluid to be compressed to flow therethrough, each of the impeller passages having one fluid inflow end. Each impeller passage has a first cross section at the fluid inflow end and a second cross section at the fluid outflow end. The compressor impeller according to the present invention is excellent in that the ratio of the size of the second cross section to the first cross section is smaller than 0.7.

  In accordance with the present invention, in a conventional compressor impeller, the ratio of the size of the second cross section to the first cross section is approximately 0,75, i.e., in any case greater than 0,7. It was.

  Drive characteristics of a centrifugal compressor equipped with a compressor impeller according to the present invention by a configuration according to the present invention having a fluid inflow end and a fluid outflow end having an area ratio or size ratio smaller than 0.7 The line in the characteristic map that characterizes the highest efficiency of such a centrifugal compressor is approximately coincident with, or nearly equal to, the line representing the characteristic curve of an internal combustion engine that is operatively connected to the centrifugal compressor. It extends in parallel and very close to it.

  Thereby, the target conflict between the compressor design and the turbocharging of the internal combustion engine is remarkably eliminated. In other words, the excessive dependence of the supply air pressure on the flow rate appears in a less pronounced manner, so that the centrifugal compressor can be operated in a state close to the optimum design according to the characteristic curve of the internal combustion engine.

  The solution according to the invention can result in a compressor characteristic map with a decreasing flow rate, especially in the partial load region, with the prudent to be able to suppress the theoretical maximum pressure, however, the actual supply pressure available And a clear increase in compressor efficiency is achieved. This is achieved in particular by the design according to the invention in the ratio of the cross section of the impeller passage at the fluid outflow end to the cross section of the impeller passage at the fluid inflow end.

  Centrifugal compressors equipped with a compressor impeller according to the invention have a clear and broader stable working area in the characteristic map, which works with reduced flow rates at part load conditions, The result is that the engine and the characteristic curve of the internal combustion engine operatively connected to each other are almost best combined as much as possible.

  According to the embodiment of the compressor impeller according to the present invention, the ratio of the size of the second cross section to the first cross section is smaller than 0.665.

  With such a configuration of the compressor impeller according to the invention, the supply pressure of the internal combustion engine can be realized up to 6 bar with essentially no loss in efficiency at the drive point.

  According to an embodiment of the compressor impeller according to the invention, the size of the first cross section is at least 1,54 times the size of the second cross section.

  With such a configuration of the compressor impeller according to the invention, the supply pressure of the internal combustion engine can be realized up to 6 bar with essentially no loss in efficiency at the drive point.

  According to a further embodiment of the compressor impeller according to the invention, the ratio of the size of the second cross section to the first cross section is less than 0,6.

  With such a configuration of the compressor impeller according to the invention, the supply pressure of the internal combustion engine can be realized up to 7 bar, essentially without any loss in efficiency at the drive point.

  According to an embodiment of the compressor impeller according to the present invention, the size of the first cross section is at least 1,67 times the size of the second cross section.

  With such a configuration of the compressor impeller according to the invention, the supply pressure of the internal combustion engine can be realized up to 7 bar, essentially without any loss in efficiency at the drive point.

  Further, according to a further embodiment of the compressor impeller according to the present invention, the compressor impeller further comprises an impeller hub having an outer periphery and a plurality of impeller blades, the impeller blades being distributed along the outer periphery of the impeller hub. The blade is provided on the impeller hub, and includes two blade side surfaces and a radially outer edge portion provided between the blade side surfaces.

  According to this embodiment of the compressor impeller according to the present invention, the outer edges of the impeller blades jointly define the outer periphery of the impeller blade, and each impeller passage is formed between adjacent impeller blades. Yes. In this case, the impeller passage is defined by the outer periphery of the impeller hub, the blade side surfaces of the adjacent impeller blades facing each other, and the outer periphery of the impeller blade. According to this embodiment of the present invention, the fluid inflow end of each impeller passage is provided radially inward, and the fluid outflow end of each impeller passage is provided radially outward.

  Furthermore, according to the embodiment of the compressor impeller according to the present invention, the impeller passage is divided between the fluid inflow end and the fluid outflow end and the impeller passage is divided into two partial passages downstream of the fluid inflow end. Each separating wall extending along the impeller passage from a separation point having a predetermined spacing to the fluid inflow end to the fluid outflow end so that the impeller passage is the only fluid And an inlet and two fluid outlets juxtaposed in the circumferential direction of the compressor impeller.

  In such a separating wall, the separating wall is compressed on the one hand without reducing the cross section of the respective impeller passage at the fluid inlet end, and on the other hand the mechanical work or kinetic energy performed by the compressor impeller. There is the advantage of better transmission to the fluid to be.

  According to an embodiment of the compressor impeller according to the present invention, each of the two partial passages of each impeller passage at the fluid outflow end has one outflow cross section, and each outflow cross section of each of the two partial passages. Is the same as the size of the second cross section.

  This form of the invention provides additional flexibility or freedom of configuration for optimal design of the compressor impeller. This is because the two outflow cross sections of the two partial passages of each impeller passage can be formed with the same size or with different sizes as required.

  Preferably, the separation wall is formed by each auxiliary blade, and the auxiliary blade corresponds to the impeller blade in its shape and radial extension, and corresponds to a predetermined interval on the fluid inflow end side. It is formed to be shorter than the impeller blade by the dimension to be adjusted.

  This form of the compressor impeller according to the invention assists in transferring the mechanical work or kinetic energy performed by the compressor impeller to the fluid to be compressed in a particularly advantageous manner.

  According to a second aspect of the present invention, a centrifugal compressor for a turbocharger, the centrifugal compressor comprising a fluid inlet for receiving a main flow of fluid to be compressed in the centrifugal compressor; A compressor housing having a feedback passage and a compressor impeller according to one or more or all of the aforementioned embodiments of the present invention, the compressor impeller downstream of the fluid inlet in the mainstream flow direction. The compressor housing is rotatably supported by the compressor housing, and the feedback passage radially surrounds the compressor impeller from the first inner periphery of the compressor housing at the fluid inlet. A centrifugal compressor is provided that extends to two inner peripheries so that a side stream of the fluid to be compressed can be formed along the feedback path.

  The configuration according to the present invention of the fluid inflow end and the fluid outflow end of the compressor impeller of the centrifugal compressor according to the present invention improves the drive characteristics of the centrifugal compressor and characterizes the highest efficiency of the centrifugal compressor The lines in the characteristic map extend approximately coincident with or approximately parallel to the line representing the characteristic curve of the internal combustion engine operatively connected to the centrifugal compressor.

  Thereby, the target conflict between the design of the centrifugal compressor and the turbocharging of the internal combustion engine is greatly eliminated. In other words, the excessive dependence of the supply air pressure on the flow rate appears in a less pronounced manner, so that the centrifugal compressor can be operated in a state close to the optimum design according to the characteristic curve of the internal combustion engine. Moreover, the solution provided according to the invention can eliminate the blow-off flap of the centrifugal compressor in some cases.

  The solution according to the invention can result in a centrifugal compressor characteristic map with a decreasing flow rate, especially in the partial load region, with the acceptability that the theoretical maximum pressure is suppressed, however, the supply that can actually be procured. A clear increase in atmospheric pressure and centrifugal compressor efficiency is achieved. This is achieved in particular by the design according to the invention in the ratio of the cross section of the impeller passage at the fluid outflow end to the cross section of the impeller passage at the fluid inflow end of the compressor impeller.

  The centrifugal compressor according to the present invention has a clear and broader stable working area in the characteristic map, which works with the reduced flow rate at part load conditions, and the centrifugal compressor and its working connection The result is that the characteristic curves of the internal combustion engines that have been made fit together almost as best as possible.

  The compressor housing of the centrifugal compressor according to the present invention is provided with a feedback passage allowing recirculation inside the compressor, so that between the achieved pressure ratio or total pressure ratio and the flow rate or volume flow rate. A very wide range of proportions along the pump limit is supported.

  In other words, the form of the recirculation chamber or feedback passage is the pump limit, and the flow of the fluid removed from the outer periphery of the impeller blade (or the return of the flow of the fluid as a side flow opposite to the main flow direction). , Upstream of the compressor impeller, fed back to or with the mainstream and recombined, which ensures that a flow rate stabilization in the compressor impeller can occur.

  On the other hand, the feedback passage is a secondary part or side stream of the transfer medium or the fluid to be compressed, which flows through the feedback passage in the mainstream flow direction, especially when the compressor speed is high. It may be designed to rejoin the main flow of fluid downstream of the critical first cross section.

  That is, the calculation and placement of the feedback path can concentrate stable compressor drive support at the pump limit or shut-off limit, but any variation or step between these two extreme embodiments. There can be.

  Preferably, for this purpose, according to the embodiment of the centrifugal compressor according to the invention, the feedback passage is provided with guide blades in order to influence the direction and / or amount of flow of the secondary flow.

  Preferably, the feedback passage includes a first end portion at the fluid supply inlet and a second end portion in the vicinity of the fluid inflow end portion of the impeller passage.

  According to the embodiment of the centrifugal compressor according to the present invention, the centrifugal compressor is formed as a single-stage centrifugal compressor.

  According to a third aspect of the present invention, there is provided an exhaust gas turbocharger for an internal combustion engine comprising an exhaust gas turbine and a centrifugal compressor according to one or more or all of the aforementioned embodiments of the present invention.

  A vehicle combustion engine (internal combustion engine) equipped with such an exhaust gas turbocharger has particularly high performance utilization and lower fuel consumption.

  In the following, the invention will be described in more detail on the basis of preferred embodiments and in connection with the attached figures.

It is a characteristic map of the working area of a normal centrifugal compressor. 1 is a schematic view of a centrifugal compressor according to an embodiment of the present invention. It is the schematic of the compressor impeller of the centrifugal compressor of FIG. It is a characteristic map of the working area of the centrifugal compressor of FIG. FIG. 5 is a diagram in which the characteristic map of FIG. 1 and the characteristic map of FIG. 4 are overlaid for comparison purposes.

  In the following, an exhaust gas turbocharger of an internal combustion engine (not fully shown) according to an embodiment of the invention will be described in connection with the attached FIGS.

  An exhaust gas turbocharger according to the present invention is formed as a diesel engine, an exhaust gas turbine (not shown) connected on the inlet side to an exhaust gas system of an internal combustion engine (not shown) of an automobile (not shown), And a single-stage centrifugal compressor 1 (shown in FIGS. 2 and 3) connected to an exhaust gas turbine so as to be rotationally driven through a drive shaft (not shown).

  Centrifugal compressor 1 has a fluid inlet 11 for receiving the mainstream H of atmospheric outside air to be compressed and possibly filtered in centrifugal compressor 1, and for discharging the compressed outside air A compression housing 10 having a fluid discharge port (not shown), a fluid discharge port fluidly connected to the air intake port of the internal combustion engine, and a feedback passage 12 is provided.

  Further, the centrifugal compressor 1 is provided downstream of the fluid inlet 11 in the flow direction of the main flow H, indicated by the tip of a line arrow indicating the main flow H as a symbol, and is rotatable to the compressor housing 10. A compressor impeller 20 is provided.

  The feedback passage 12 extends from the first inner peripheral portion 13 of the compressor housing 10 located at the fluid inlet 11 to the second inner peripheral portion 14 of the compressor housing 10 that radially surrounds the compressor impeller 20. As a result, the secondary flow N of the outside air to be compressed along the feedback passage 12 may occur in the opposite direction or the same direction as the flow direction of the main flow H depending on the driving conditions.

  The feedback passage 12 is formed by a ring-shaped cavity 15 on the inner periphery of the compressor housing 10 and a ring element 16 attached to the fluid inlet 11. As a result, the feedback passage 12 is formed by the fluid inlet 12. 11 and a second end 12b located in the vicinity of the fluid inflow end 23a of each impeller passage 23 of the compressor impeller 20.

  The feedback passage 12 is provided with a fixed or adjustable guide blade 17 which influences the flow direction and / or flow rate of the side stream N.

  For example, Patent Document 1 discloses a feedback passage corresponding to the feedback passage 12 with respect to its action.

  The compressor impeller 20 includes an impeller hub 21 that is non-rotatably fixed to the drive shaft and includes an outer periphery 21 a and a plurality of impeller blades 22. The impeller blade is disposed on the outer periphery 21 a of the impeller hub 21. Are provided in the impeller hub 21 in a state of being evenly distributed along the circumferential direction along the blade side surfaces 22a and 22b extending in two lateral directions, and extending between the blade side surfaces 22a and 22b. And a radially outer edge 22c.

  The outer periphery (considered as a rotating body) of the impeller blades 22 is defined by the cooperation of the radially outer edges 22c of the respective impeller blades 22 and is compressed between the adjacent impeller blades 22 and 22, respectively. Each impeller passage 23 is formed to allow the outside air (fluid) to flow through.

  The impeller passages 23 formed in this way are each provided with one fluid inflow end portion 23a located radially inward (provided in the vicinity of the drive shaft) and radially outward (fluid inflow end portion 23a). In comparison, a fluid outflow end 23b (separated from the drive shaft in the radial direction) is provided. The outer periphery of the impeller blade 22 is surrounded by the second inner peripheral portion 14 and the ring element 16 of the compressor housing 10 from the outside in the radial direction with a slight gap therebetween.

  The impeller passages 23 respectively include an outer periphery 21a of the impeller hub 21, blade side surfaces 22a and 22b of the adjacent impeller blades 22 and 22 facing each other, an outer periphery of the impeller blade 22 or a second inner peripheral portion 14 of the compressor housing 10 and a ring. Defined by the element 16.

  Each of the impeller passages 23 includes one separation wall having the shape of the auxiliary blade 24 between the fluid inflow end portion 23a of each of the impeller passages 23 and the fluid outflow end portion 23b of each of the impeller passages 23. In the extension in the radial direction, it coincides with the impeller blade 22, but the fluid inflow end portion 23a is formed shorter than the impeller blade 22 by a predetermined interval.

  In other words, each impeller passage 23 is divided into two partial passages 23c and 23d downstream of the fluid inflow end portion 23a, and the auxiliary blade 24 functioning as a separation wall is predetermined with respect to the fluid inflow end portion 23a. Extends along the impeller passage 23 from the separation point T (shown only in FIG. 3) to the fluid outflow end 23b, so that the impeller passage 23 is the only fluid inlet and And two fluid outlets juxtaposed in the circumferential direction of the compressor impeller 20.

  Each impeller passage 23 has a first cross section, that is, an inflow cross section AE at the fluid inflow end portion 23a.

Two partial passage 23c of the impeller passages 23, 23d, respectively, in the fluid outlet end 23b of the corresponding impeller passageway 23, and a single outflow cross section AA ₁ or outflow cross section AA _2. In this embodiment, two equal sized outflow cross sections AA ₁ , AA ₂ of the two partial passages 23 c, 23 d together form a second cross section or the total outflow cross section AA of the respective impeller passage 23. (AA = AA ₁ + AA ₂ ). In a further embodiment of the invention, the two outflow cross sections AA ₁ , AA ₂ may be of different sizes.

  In the embodiment of the present invention, the ratio GV of the size of the second cross section AA to the first cross section AE is defined by GV = AA / AE and is smaller than 0.7.

  If the size ratio GV is smaller than 0.7, the supply pressure of the internal combustion engine can be 5.5 bar or more without any loss in efficiency at the drive point.

  In a further embodiment of the invention, the ratio GV of the size of the second cross section AA to the first cross section AE is less than 0.65.

  If the size ratio GV is smaller than 0.65, the supply pressure of the internal combustion engine can be increased to 6 bar without any loss in efficiency at the drive point.

  In a still further embodiment of the invention, the ratio GV of the size of the second cross section AA to the first cross section AE is less than 0.6.

  If the size ratio GV is smaller than 0.6, the supply pressure of the internal combustion engine can be increased to 7 bar without any loss in efficiency at the drive point.

  FIG. 4 shows a characteristic map of the work area of the centrifugal compressor 1 according to the present invention. FIG. 5 is a diagram in which the characteristic map of FIG. 1 and the characteristic map of FIG. 4 are superimposed to compare the characteristic map of FIG. 1 with the characteristic map of FIG.

As is clear from FIGS. 4 and 5 (in connection with FIG. 1), the sizes of the fluid inflow end 23a and the fluid outflow end 23b of the compressor impeller 20 of the centrifugal compressor 1 according to the present invention are as follows. The line WG ′ _opt in the characteristic map of the ratio, that is, the area ratio GV, indicates the highest efficiency of the centrifugal compressor 1 due to the improvement of the driving characteristics of the centrifugal compressor 1 by the configuration according to the present invention. The line extends along the line BC ′ representing the characteristic curve of the internal combustion engine that is functionally connected to the compressor 1, or substantially parallel to the line BC ′ and very close to the line BC ′.

Thereby, the target conflict between the design of the centrifugal compressor and the turbocharging of the internal combustion engine, known from the prior art, is very greatly eliminated. In other words, the excessive dependence of the supply pressure or the total pressure ratio Π _{tot on} the flow rate or volume flow rate V is exposed in a less pronounced manner, so that the optimum design according to the characteristic curve BC ′ of the internal combustion engine The centrifugal compressor 1 can be operated in a state close to.

The centrifugal compressor 1 according to the present invention has a characteristic map showing that the flow rate, that is, the volume flow rate V is decreased particularly in the partial load region. In this characteristic map, the theoretical maximum pressure (lines WG _opt , WG The uppermost end of ' _opt ' is suppressed, but an increase ΔΔ _tot between the supply pressure actually available and the efficiency of the centrifugal compressor 1 is remarkably achieved.

  The centrifugal compressor 1 according to the invention has a clear and broader stable working area in the characteristic map, which works with a reduced flow rate or volumetric flow rate V under partial load conditions, And the characteristic curve BC ′ of the internal combustion engine which is operatively connected to this result in the best possible combination with each other.

The compressor housing 10 of the centrifugal compressor 1 according to the present invention is provided with a feedback passage 12 that allows recirculation inside the compressor, so that the pressure ratio or total pressure ratio Π _tot achieved and the flow rate or volume. A very wide range of proportions with the flow rate V along the pump limit PG ′ is supported.

  The configuration of the guide blade 17 and the size / arrangement of the feedback passage 12 are the pump limit PG ′, and the flow of the fluid removed from the outer periphery of the impeller blade 22 and the auxiliary blade 24 (or opposite to the flow direction of the main flow H). Of the fluid flow as a secondary flow N) is fed back to the main flow H or upstream with the main flow H upstream of the compressor impeller 20 to recombine, thereby stabilizing the flow ratio in the compressor impeller 20. It is guaranteed that it can occur.

  On the other hand, the feedback passage 12 is a compression that flows through the feedback passage 12 in the flow direction of the main flow H, particularly when the rotation speed of the compressor is high, depending on the configuration of the guide blade 17 and the size / arrangement of the feedback passage 12. It may be arranged such that the substream N of the outside air to be recombined with the main flow H of the outside air to be compressed downstream of the first cross section AE that is decisive for the flow or volume flow V.

  That is, the structure of the feedback passage 12 can concentrate stable compressor drive support on the pump limit PG ′ or the shut-off limit SG ′, but any variation between these two extreme embodiments is possible. Or there can be stages.

DESCRIPTION OF SYMBOLS 1 Centrifugal compressor 10 Compressor housing 11 Fluid inlet 12 Feedback path 12a 1st end part 12b 2nd end part 13 1st inner peripheral part 14 2nd inner peripheral part 15 Cavity 16 Ring element 17 Guide blade 20 Compressor impeller 21 Impeller hub 21a Outer periphery 22 Impeller blade 22a Blade side surface 22b Blade side surface 22c Radial outer edge 23 Impeller passage 23a Fluid inflow end portion 23b Fluid outflow end portion 23c Partial passage 23d Partial passage 24 Auxiliary blade T Separation point AE First cross section surface)
AA 2nd cross section (total outflow cross section)
AA ₁ Outflow cross section AA ₂ Outflow cross section H Main flow N Subflow Π _tot Total pressure ratio ΔΠ _tot increase V Volume flow PG Pump limit SG Cutoff limit WG _opt Optimal efficiency BC Internal combustion engine characteristic curve n _konst constant rotation speed PG 'Pump limit SG' cutoff limit WG ' _opt optimum efficiency BC' characteristic curve of internal combustion engine n ' _konst constant rotation speed

Claims (11)

A compressor impeller (20) for a centrifugal compressor (1) having a plurality of impeller passages (23) for allowing the fluid to be compressed to flow through it,
Each of the impeller passages (23) includes one fluid inflow end (23a) and a fluid outflow end (23b), and each of the impeller passages (23) has the fluid inflow end (23a). ) Having a first cross section (AE) and a second cross section (AA) at the fluid outflow end (23b), in the compressor impeller (20),
The compressor impeller (20), wherein a ratio (GV) of a size of the second cross section (AA) to the first cross section (AE) is smaller than 0.7.   The compressor impeller (20) according to claim 1, wherein the ratio (GV) of the size of the second cross section (AA) to the first cross section (AE) is less than 0.65. .   The compressor impeller according to claim 1 or 2, wherein the ratio (GV) of the size of the second cross section (AA) to the first cross section (AE) is smaller than 0.6. 20). The compressor impeller (20)
An impeller hub (21) having an outer periphery (21a);
A plurality of impeller blades (22) provided on the impeller hub (21) in a state of being distributed along the outer periphery (21a) of the impeller hub (21), each of the impeller blades (22) being A plurality of the impeller blades (22) including two blade side surfaces (22a, 22b) and a radially outer edge (22c) provided between the blade side surfaces (22a, 22b);
With
The outer edge (22c) of the impeller blade (22) cooperates to form the outer periphery of the impeller blade (22),
Each of the impeller passages (23) is formed between adjacent impeller blades (22, 22), and each of the impeller passages (23) is adjacent to the outer periphery (21a) of the impeller hub (21). The impeller blades (22, 22) are formed by the blade side surfaces (22a, 22b) facing each other and the outer periphery of the impeller blade (22),
The fluid inflow end portion (23a) of each of the impeller passages (23) is provided on the radially inner side, and the fluid outflow end portion (23b) of each of the impeller passages (23) is provided on the radially outer side. The compressor impeller (20) according to any one of claims 1 to 3, wherein the compressor impeller (20) is provided. Each of the impeller passages (23) has one separating wall dividing the impeller passage (23) into two partial passages (23c, 23d) downstream of the fluid inflow end portion (23a). Between the end (23a) and the fluid outflow end (23b),
The separation wall extends along the impeller passage (23) from the separation point (T) having a predetermined interval to the fluid inflow end (23a) to the fluid outflow end (23b). As a result, the impeller passage (23) is provided with a single fluid inlet and two fluid outlets juxtaposed with each other in the circumferential direction of the compressor impeller (20). The compressor impeller (20) according to any one of claims 1 to 4, characterized in that. Wherein said impeller passages (23) each of two of said subchannels (23c, 23d) of the fluid outlet end (23b) is, and respectively provided with one outflow cross section _(AA 1, AA _2),
The sum of the sizes of the outflow cross sections (AA ₁ , AA ₂ ) of the _two partial passages (23c, 23d) is the same as the size of the second cross section (AA). The compressor impeller (20) according to claim 5. The separating wall is formed by a respective auxiliary blade (24);
The auxiliary blade (24) corresponds to the impeller blade (22) in the radial extension of the auxiliary blade (24), and on the fluid inflow end side, the dimension corresponding to a predetermined interval is provided. Compressor impeller (20) according to claim 5 or 6, characterized in that it is shorter than the impeller blade (22). A centrifugal compressor (1) for a turbocharger,
A compressor housing (10) having a fluid inlet (11) and a feedback passage (12) for receiving a main flow (H) of fluid to be compressed in the centrifugal compressor (1);
8. The compressor according to claim 1, which is provided downstream of the fluid inlet (11) in the flow direction of the main flow (H) and is rotatably supported by the compressor housing (10). A compressor impeller (20) according to claim 1,
In the centrifugal compressor (1) comprising:
The feedback passage (12) surrounds the compressor impeller (20) in the radial direction from the first inner peripheral portion (13) of the compressor housing (10) located at the fluid supply inlet (11). The compressor housing (10) extends to the second inner peripheral portion (14), whereby a substream (N) of the fluid to be compressed flows along the feedback passage (12). A centrifugal compressor (1) characterized by the above.   9. A guide blade (17) according to claim 8, characterized in that the feedback passage (12) is provided with a guide blade (17) in order to influence the flow direction and / or the flow rate of the substream (N). Centrifugal compressor (1).   The feedback passage (12) has a first end (12a) located at the fluid inlet (11) and a second end located near the fluid inflow end (23a) of the impeller passage (23). The centrifugal compressor (1) according to claim 8 or 9, comprising a section (12b).   An exhaust gas turbocharger for an internal combustion engine comprising an exhaust gas turbine and the centrifugal compressor (1) according to any one of claims 8 to 10.

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