Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/40—Casings; Connections of working fluid
  • F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
  • F04D29/44—Fluid-guiding means, e.g. diffusers
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/26—Rotors specially for elastic fluids
  • F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
  • F04D29/30—Vanes
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/26—Rotors specially for elastic fluids
  • F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
  • F04D29/284—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/40—Casings; Connections of working fluid
  • F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
  • F04D29/4206—Casings; Connections of working fluid for radial or helico-centrifugal pumps especially adapted for elastic fluid pumps
  • F04D29/4213—Casings; Connections of working fluid for radial or helico-centrifugal pumps especially adapted for elastic fluid pumps suction ports
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D29/00—Details, component parts, or accessories
  • F04D29/66—Combating cavitation, whirls, noise, vibration or the like; Balancing
  • F04D29/68—Combating cavitation, whirls, noise, vibration or the like; Balancing by influencing boundary layers
  • F04D29/681—Combating cavitation, whirls, noise, vibration or the like; Balancing by influencing boundary layers especially adapted for elastic fluid pumps
  • F04D29/685—Inducing localised fluid recirculation in the stator-rotor interface

Definitions

• 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 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.
• Fig. 1 illustrates, showing a total pressure ratio ⁇ to t via
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• the size ratio of the second cross-sectional area to the first cross-sectional area is smaller than 0.65.
• a size of the first cross-sectional area is at least 1.54 times a size of the second cross-sectional area.
• 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.
• the size ratio of the second cross-sectional area to the first cross-sectional area is smaller than 0.6.
• boost pressures to the internal combustion engine of up to 7 bar can be realized substantially without sacrificing the efficiency at the operating point.
• a size of the first cross-sectional area is at least 1.67 times a size of the second cross-sectional area.
• boost pressures to the internal combustion engine of up to 7 bar can be realized substantially without sacrificing the efficiency at the operating point.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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 fluideinthend herb 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.
• a radial compressor for a turbocharger 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• guide vanes are preferably arranged for influencing a flow direction and / or a flow rate of the secondary flow.
• 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.
• this is designed as a single-stage centrifugal compressor.
• 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.
• 1 shows a characteristic diagram of the working range of a conventional radial compressor.
• FIG. 2 shows a schematic view of a radial compressor according to an embodiment of the invention.
• FIG. 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.
• the exhaust gas turbocharger 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 lubanthebsverbunden 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.
• fixed or adjustable guide vanes 17 are arranged in the feedback passage 12.
• 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.
• 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.
• 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 two outlet cross-sectional areas AA-i, AA 2 can also be of different sizes.
• 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:
• the size ratio GV of the second cross-sectional area AA to the first cross-sectional area AE is smaller than 0.65.
• 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.
• the size ratio GV of the second cross-sectional area AA to the first cross-sectional area AE is less than 0.6.
• FIG. 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.
• 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 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.
• the radial compressor 1 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
• the radial compressor 1 according to the invention has an apparent in the map wider stable working area, resulting in interaction with
• V 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 compressor housing 10 of the radial compressor 1 has the feedback passage 12, which allows a compressor internal recirculation, a maximum proportionality between the achieved pressure ratio or
• 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.
• 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

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Citations (5)

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• 2009
  • 2009-06-08 DE DE102009024568A patent/DE102009024568A1/de not_active Withdrawn
• 2010
  • 2010-01-19 CN CN201080025586.7A patent/CN102803739B/zh not_active Expired - Fee Related
  • 2010-01-19 KR KR1020127000160A patent/KR101369601B1/ko active IP Right Grant
  • 2010-01-19 EP EP10706906A patent/EP2440791A1/de not_active Withdrawn
  • 2010-01-19 JP JP2012514346A patent/JP2012529585A/ja active Pending
  • 2010-01-19 WO PCT/DE2010/050001 patent/WO2010142287A1/de active Application Filing

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