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The Development of a Centrifugal Compressor ImpellerC. Xu a & R. S. Amano aa University of Wisconsin-Milwaukee, Milwaukee, WI, USAPublished online: 11 Jun 2009.

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International Journal for Computational Methods in Engineering Science and Mechanics, 10:290–301, 2009Copyright c© Taylor & Francis Group, LLCISSN: 1550–2287 print / 1550–2295 onlineDOI: 10.1080/15502280903023165

The Development of a Centrifugal Compressor Impeller

C. Xu and R. S. AmanoUniversity of Wisconsin-Milwaukee, Milwaukee, WI, USA

An impeller is one of the key components of industrial centrifu-gal compressors and turbochargers. Aerodynamic and structuredesigns of the impeller are critical to the success of the whole com-pressor stages. The requirements for efficiency and operating rangeof industrial centrifugal compressors and turbochargers have beenincreased dramatically compared with the situation in the past.The efficiency of a newly developed, low-pressure-ratio centrifugalcompressor has reached the possible level of the machine. How-ever, the efficiency level of an intermediate- and high-pressureratio machine still has gaps between the current state-of-the-artand possible level. The challenge for centrifugal compressor de-sign is to keep the efficiency level at state-of-the-art and increasethe compressor operating range. Increase of the compressor oper-ating range without sacrificing compressor peak efficiency is dif-ficult to achieve. The product globalization requires one productdesign, which can be used in all locations. In some counties, due tothe technology differences, electricity frequency variations couldbe 10%. Turbocharger compressors work at different rotationalspeeds for the majority of the time. The compressor impeller rotat-ing speeds change in certain ranges. The impeller rotating speedvariation makes the impeller structure design more challenging.In this study, a full-3D impeller was designed to optimize impelleraerodynamic performance and structure characteristics.

Keywords Impeller, CFD, FEA

INTRODUCTIONCentrifugal compressors have a wide application in indus-

trial, aerospace, and automobile industries. In a centrifugal com-pressor the flow enters the compressor axially and then turns inthe radial direction out from the impeller. The flow then enters aradial annular vaned or vaneless diffuser. The flow exit from thediffuser needs a volute or collector to deliver the flow to the nextstage or send it to the next components [1–3]. The basic com-ponents of the radial compressor are shown in Fig. 1. Unlike anaxial compressor or fan [4, 5], the work input for a centrifugalcompressor is almost independent of the nature of the flow. Acentrifugal compressor can be designed with much higher De

Address of correspondence to R. S. Amano, University ofWisconsin-Milwaukee, Milwaukee, WI 53201, USA. E-mail: amano@uwm.edu

Haller number than an axial compressor can achieve. So it is pos-sible for a centrifugal compressor to have a much higher stagepressure ratio than axial ones. Centrifugal compressors havewide applications for a mass flow rate less than 10 kg/s [1, 5].

The small flow centrifugal compressors have wide applica-tions in turboshaft aircraft engines, turbochargers, and industry.However, being in a low-power class, these compressors needto be inexpensive to manufacture and operate, requiring that thecompressor be a simple design having fewer parts and relativelylarge tolerance. Moreover, the five-axis machine is now a com-mon tool for impeller machining, but most other parts shouldbe fabricated by using other types of machining to reduce themanufacturing costs.

The turbomachinery industry is increasingly interested in us-ing optimization procedures that enable enhanced compressorefficiency and widen operating ranges. During the compres-sor development, designers always compromise between a peakattainable efficiency and an overall operating range [6]. Com-pressor design normally starts with a meanline program at eachindividual operating point on a map, then throughflow calcula-tion is performed, and finally the impeller, diffuser, and voluteare designed. In this study, a recently developed turbomachineryviscous optimal method [7, 8] for axial machines was furtherextended to a centrifugal compressor design [9]. The main focusof this study lies in the development of a small flow compressorwhere flow coefficient φ = Q/N/D3 = 0.145. Q, N, and D arethe volume metric flow (m3/s), rotational speed (rps), and im-peller tip diameter (m). The design pressure ratio and the flowrate are 3.65 and 0.75 kg/s, respectively, at the design condition.The total to static efficiency required to be higher than 84% andthe stability operating range (SB) great than 38%. The com-pressor design employs a global optimization viscous processfor achieving efficiency and stability targets [8]. Good surgemargins were achieved without use of a variable geometry for astable operation. Special attention has been paid to the tip clear-ance profile to permit large clearance without too much lostefficiency during the impeller design. The compressor devel-oped in this study consists of three major parts: an impeller, lowsolidity diffuser, and volute. In this study, particular attentionwas paid mainly to impeller design and analyses.

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FIG. 1. Centrifugal compressor.

A variable speed centrifugal compressor impeller with a com-pound lean blade was developed [9, 10]. The newly developedimpeller defines a blade pressure side that extends from theleading edge to the trailing edge and is convex near the leadingedge and flat near the trailing edge. The impeller was leaned op-posite to rotation at the inlet portion and leaned back at exit. Theclearance between impeller and casing is non-uniform, definedby a high-order equation. The flow and structure analyses wereperformed. Some basic performance tests and structure modaltests were performed. The analyses agreed with test results.

EXPERIMENTAL STUDYTwo impeller designs were installed with the same diffuser

and scroll stage. The experimental studies were performed inthe Centrifugal Compressor Development Laboratory. The per-former test and data acquisition procedures were based on theASME performance test procedure [11]. The test compressorconsists of: inlet pipe, impeller, vaneless diffuser, low solidityvaned diffuser, vaneless diffuser, and volute as shown in Fig. 1.The test circuit is an open loop testing system as shown in Fig.2. The flow rate is regulated by a butterfly valve located at thedischarge pipe of the compressor. The main dimensions of thetest compressor are listed in Table 1. Reference [9] providesmore details for the test compressor.

The test uncertainty for the flow rate is less than 2% at 95%confidence level. The head and efficiency uncertainties werekept under 2.2% and 2.5%, respectively, with the same confi-dence level based on the system uncertainty analyses [9]. Theperformance characteristic was tested through five total temper-ature and total pressure measurements at the inlet of compressorand discharge of the volute exit cone for both impellers. Due tothe motor system problems after running the traditional impellerperformance test, the off-design point test for a new full-three-dimensional impeller was not able to performance. Only design

TABLE 1Main design information of the test compressor

Mass flow rate mShaft speed N = 60000 rpm = 0.77 kg/s

Impeller outlet radius R2 =82.90 mm

Impeller outlet width b2 =7.85 mm

Out blade angle β2 = 55◦ Design pressure ratio π =3.60

Impeller blade number Zi =19

Numbers of diffuser vanesZd = 9

Diffuser inlet radius R3 =88.70 mm

Volute inlet radius R5 =170.00 mm

point performance was tested for a full-three-dimensional im-peller. More testing for off-design point performance will beconducted after fixing the motor problem. For determining thestatic pressure ratio, five pressure transducer taps were mountedon the circumferentially distributed wall at discharge of voluteexit cone, which were used to measure the pressure. The massflow was measured by using an ASME [10] nozzle located atthe end of the discharge pipe. The tip clearance of the impellerat tip was measured with an alumina pin. The design tip run-ning clearances at inlet and exit are about 0.2 mm and 0.1 mm,respectively. The performance test results are discussed in theCFD calculation section.

For comparing the structure analyses with the experimentalresults, vibration tests were performed for the full 3D impellerutilizing a LDV (Laser Doppler Velocitmeter) and an impacthammer. A calibrated impact hammer was used to excite the im-peller at different points and the resulting frequency and modeshape were measured by fast Fourier transform (FFT) and cir-cular sampling of the part. Post processing yields a spectralresponse pattern and mode shape. The frequency test results arereported in the following section.

FIG. 2. Test rig of a single stage compressor.

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FIG. 3. FEA boundary conditions and meshed for different impeller blades.

FEA AND MODAL ANALYSESIn a modern centrifugal compressor, due to increased re-

quirements of compressor performance, the blade thickness hasa trend to reduce. This can be achieved by the selection of bettermaterial and conducting more detailed finite-element analysis(FEA).

Finite element analysis is a powerful tool to help find correc-tion factors that make more cost-effective, simplified analysesas accurate as possible [12]. Stress analyses of the centrifugalimpeller wheel during the design stage helps diagnose possibledesign problems and avoid failure. The finite element methodcan identify high-stress locations as well as vibrations in thecentrifugal compressor impellers. Integration optimization be-tween FEA and CFD is more popular in industry for turboma-chinery designs [7,8]. Optimizations guided design changes thatimproved the integrity of the compressor.

In this study, a preprocess BLADEPROTM [13] was used forAnsys [14] solver. The preprocess allows the user to specify Ra-dial (R), Tangential (T), and Axial (A) boundary conditions forcover-face nodes for both static (single-blade steady stress) anddynamic (modal) analyses as shown in Fig. 3. A.B.C., R.B.C.,and R.A.B.C. in Fig. 3 represent Axial constrain boundary con-dition, Radial constrain boundary condition and Radial and Ax-ial constrain boundary conditions. The temperature and pressureprofiles on the blade were calculated based on the CFD analyses.After generating the mesh and setting boundary conditions, thestress and modal analyses were performed. The material prop-erties such as Youngs Modulus, material density and Poisson’sratio and thermal properties such as coefficient of thermal ex-pansion were selected from BLADEPROTM material database(Ti-6A1-4v Titanium Alloy). Calculations were performed on aDell Precision 490 personal computer. The calculation time forstress analysis was about 30 seconds.

The calculation showed that overall von Mises stress is sim-ilar for both full three-dimensional blade and traditional three-dimensional design, those of which are the cases without bowand compound lean features. The von Mises stress contours nearthe impeller bore area were enlarged and are shown in Fig. 4.The max stress near bore area is about 82.4 ksi for the fully 3Dblade and is about 82.3 ksi for traditional blade. It is shown thatthe bore stress does not change too much for both designs. Thevon Mises stress contours near the root of the trailing edge isshown in Fig. 5. As shown in Fig. 5, the maximum von Misesstress for the full 3D blade is a little higher than the cases fortraditional blades. This is because the full 3D blade trailing edgestiffness may be smaller than a traditional blade. Less stiffnesswill cause more deflection and increase the stress. However, allthe von Mises stress level is well below the yield strength ofthe material. The maximum von Mises stress for full 3D bladepeak stress near the leading edge is about 64ksi and is about52ksi for the traditional design as shown in Fig. 6. Leading edge

FIG. 4. von Mises stress contour near bore area for different impeller blade designs.

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FIG. 5. von Mises stress contour near trailing edge for different impeller blade designs.

stress mainly depends on how much material near the leadingedge area is used. The full 3D blade design has more materialnear the leading edge area due to the compound lean and bow.Additional material increases the centrifugal force during therotation and causes higher stress. However, the maximum stressnear the root of the leading edge for both designs is far be-low yield strength. It is shown that both designs meet the stressrequirements.

After stress analyses were completed, an at-speed modalanalysis was performed in which a single-blade steady stressanalysis was first performed to include the stress stiffening ef-fects in frequency prediction. Stress stiffening effects from asingle-blade stress analyses were carried over into the frequencycalculation. At-speed modal analysis performs frequency pre-diction at a given shaft speed where both the “stress stiffening”

and “spin softening” effects are included in frequencyprediction.

A cyclic sector can be modeled using two options: the re-duced order model (ROM) and the full cyclic model. ROMuses super-elements for frequency prediction whereas the fullcyclic model uses all the elements in the cyclic sector for fre-quency prediction. The full wheel method is used when thereis a non-cyclic structure; for example, a wheel with unequalgroup lengths. The full cyclic model uses all the elements in thecyclic sector for frequency prediction. In this study, an at-speedfull cyclic modal analysis was performed on a Dell Precision490 personal computer after stress analyses. Modal analyses foreach case took about 2500 seconds.

The interference diagram is a plot of Frequency vs. NodalDiameter that is used to identify potential resonance situations

FIG. 6. von Mises stress contour near leading edge root for different impeller blade designs.

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FIG. 7. Interference diagram for different impeller blade designs.

[15]. Each mode is represented by a dot on the interferencediagram. The impulse lines in the interference diagram wereconstant speed lines corresponding to high and low limits ofthe running speeds of the machine. Intersection of this constantspeed line with one or more dots indicates potential resonanceconditions. The modal analyses results are plotted in the form ofinterference diagram as shown in Fig. 7. It can be seen that the

traditional design can not meet frequency margin requirements(5%) at mode 39 due to the excitations of nine diffuser vanes.The full 3D wheel did not change the blade hub and shroud pro-files and thickness; however, it successfully meets the frequencymargin requirements.

The frequency prediction accuracy is dependent on the modeshape and level of the frequency; simple modes like 1st bending

FIG. 8. Sketch of the computational mesh.

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FIG. 9. Convergent history of RMS mass.

and torsion can be predicted more accurately than higher com-plex modes. The full three-dimensional blade mode calculationsare more difficult than traditional three-dimensional blades. Forvalidating the calculations, the modal testing for a full 3D wheelwas conducted. It can be seen that the calculation results agreewith the measurements.

FIG. 10. Comparison of the performance curves.

CFD CALCULATIONSThe single stage compressor, from inlet pipe, impeller, vaned

diffuser, vanless diffuser and volute, is computationally ana-lyzed using a Navier-Stokes solver on a fully 3-D viscous tur-bulent flow solver CFX10.0. The geometrical discretization ofthe compressor stage is made for the computational analysis.

FIG. 11. Tip relative Mach number distribution near shroud.

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FIG. 12. Relative Mach number contour near suction side of the blade.

FIG. 13. Relative Mach number contour near pressure side of the blade.

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FIG. 14. Relative Mach number contour near mid plan between the blades.

FIG. 15. Static pressure contour near suction side.

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FIG. 16. Static pressure contour near pressure surface.

FIG. 17. Radial velocity contour near suction surface.

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FIG. 18. Radial velocity contour near pressure surface.

Structure hexahedral cells are generated to define all parts ofthe compressor as shown in Fig. 8. The meshed size for inletpipe, impeller, diffuser, and volute are 29,940, 597,902, 497,314and 768,543 nodes, respectively. The interface surface betweeninlet-impeller and impeller-diffuser is modeled by using thefrozen rotor method, i.e. the relative orientation of the two in-terface components across the interface is fixed. The calculationof the downstream surface of the interface plan was based onthe average mixing plane approach. The calculation assumed tobe converged when RMS of flow, pressure, and turbulence pa-rameters smaller than the 7th order of original errors as shownin Fig. 9.

The turbulence model is one of the factors to determine thesuccess of the CFD analyses. Different turbulence models weretested for single stage centrifugal compressor flow calculations[16]. It is shown that both k-ε two equations and zero equationturbulence models provided reasonably good results, as shownin Fig. 10. The turbulence model used in this study is based on azero equation model; only an algebraic equation was used to cal-culate the viscous contribution from turbulent eddies. The inletboundary conditions were enforced by using the total pressureand total temperature. The flow entering the pipe is assumed tobe normal to the inlet surface. The outlet boundary conditionwas applied by using a variable static pressure proportional tothe kinetic energy at the outlet. The flow rate of the compressoris changed by modifying the static pressure to kinetic energyratio at the outlet condition, which simulates different closingpositions of the butterfly valve employed in the tests. In thecalculation, no-slip boundary conditions were imposed over theimpeller blades, diffuser airfoils, and all solid walls.

The calculations were performed in an HP 8000c worksta-tion. The time steps used in the calculation were set to 5.0 ×10−5 seconds at near surge and 5.0 × 10−4 seconds at other con-ditions. The calculation was assumed to be convergent when theratio between the sum of the residuals and the sum of the fluxesfor a given variable in all the cells is reduced to at least six ordersof magnitude. Intensive grid size dependence tests were carriedout [9] and overall compressor performance was compared withdifferent mesh sizes. The final mesh size was set when the meshsize increased; the mass flow rate changed less than 0.5%.

Two types of turbulence models, standard k-ε two equationand zero equation turbulence models, were used to calculate thecompressor performance for traditional impeller. The resultsare plotted in Fig. 10. It is shown that the compressor devel-oped by this research has a wide operating range of 40% withhigh efficiency in all operating conditions. It is shown that bothturbulence models provided good predications for compressorefficiencies and head coefficients compared with the test, exceptthe compressor operated near choke. For the compressor oper-ated near choke, computations overpredicted the performancecompared with tests. For full 3D design, the calculation showsa little performance advantage, better efficiency and head at thesame flow conditions, compared with traditional impeller de-sign. However, the tests near design flow show that there areno significant benefits for both efficiency and head coefficientsnear design point. The test showed a little benefit near choke andsurge operations. To reduce the test tolerances, a few repeatedtests were performed and similar results were obtained. Uponfurther inspecting the impellers, it was found that the impellersurface finish is about 6.3 micrometers near hub and about 3.2

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FIG. 19. Relative flow angle contour at blade exit.

micrometers on the blades. Because the calculations assumedthat all walls on the impeller are a smooth wall, the frictionloss estimations may be smaller than the test impeller. When acompressor is operated at a near design condition, both full 3Dimpeller and traditional impeller have no massive separation;the friction loss for the full 3D impeller is larger than the tradi-tional blade due to bowed and lean features increasing the wetarea. Further examination of the impeller calculations at designpoint for the compounded lean 3D impeller and conventionalimpeller showed that the total to total efficiency for 3D impellerand traditional impeller were 93.33% and 92.99%, respectively.It can be seen that the gain of the efficiency for 3D design isabout 1/3%. This performance gain in the impeller becomes lesssignificant for a whole stage. If the test part surface finish is notvery high, the effect of the performance due to a 3D design willbe washed out by frictional loss.

The tip relative Mach number contours near the tip of theimpellers for both full 3D and traditional designed impellerare shown in Fig. 11. It is shown that the traditional designleading edge area had a relatively larger Mach number com-pared with the full 3D design impeller. This may indicate thatthe full 3D design had a smaller portion of the flow pass-ing through the blade tip range. It is favorable for impeller

efficiency if it shows a smaller flow in the tip high loss range.This is one of the reasons why the full 3D design shows a betterefficiency.

The relative Mach number contours near the suction surface,pressure surface, and mid-span plans are shown in Figs. 12, 13,and 14. It can be seen that the compounded lean blade has arelatively larger high Mach number area than the conventionalblade near the suction side of the blade in the middle of the flowpath. At a near inlet area the full 3D design depicts a smallerhigh Mach number area than the traditional impeller design.On the pressure side, the low Mach number region of the full3D impeller is smaller than the conventional design. Mid-spanMach number distributions also show that a full 3D design pro-vides more uniform Mach number distributions. The uniformMach number distributions would benefit the compressor per-formance.

The static pressure contours near the pressure and suctionside are shown in Figs. 15 and 16. The static pressure contoursshow that the traditional blade has less static pressure rise com-pared with a lean blade. Figures 17 and 18 show the relativeradial velocity contour near suction and the pressure surface ofthe impeller blade, respectively. It is shown that there is a largerhigh radial velocity region for the traditional impeller both near

FIG. 20. Secondary velocity contour at blade exit.

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FIG. 21. Relative Mach number contour at blade exit.

the suction and pressure side at about 50% of the flow path loca-tion of the impeller. It also shows that this higher velocity regionis also closer to impeller shroud for traditional blade impeller.It is interesting to point out that the radial velocities near theshroud at 50% flow path location drop very fast near the impellerexit. At the impeller exit, the radial velocities near shroud arevery low.

The flow angle contour at blade exit plan in Fig. 19 showsthat the compounded lean full 3D blades design provides amore uniform exit angle distribution. The larger angle flow areais smaller for the full 3D design. This characteristic will benefitthe vaned diffuser performance. This allows a 3D design to geta little more efficiency compared with a traditional design. Thesecondary flow contour at the exit of the impeller is shown inFig. 20. It is shown here that both designs give a single vortexstructure and the main secondary flow center is located near thesuction side of the blade. In the traditional design, the vortexcenter is located closer to the shroud side and also shows muchstronger secondary flows. The relative Mach number distribu-tion in Fig. 21 shows that the regions for both high and low Machnumber areas are larger for the traditional design. It indicatedthat the flow is less uniform at the impeller exit for traditionaldesign.

CONCLUSIONSIn this study, a new full 3D centrifugal impeller was designed

and analyzed for both structure and flow. Calculations showedthat a full 3D blade is both advantageous in a structure vibrationand an aerodynamic performance. The test results show thatthe full 3D impeller did not show any significant benefits inperformance near the design point due to the increase in frictionlosses on the surface of the bowed blades, whose surface areais greater than that of the conventional impeller. In cases ofnear surge and choke, full 3D impeller separation is reducedin the bowed blade impeller so that the stagnation pressureratio and efficiency are increased. Structure analyses showedthat the bowed blade could change the impeller into higherorder modes separations. Natural frequencies of the impeller had

different values compared with the traditional blade. This featureoffers benefits that previously have not been reported in theliterature.

REFERENCES1. D. Japikse, Centrifugal Compressor Design and Performance, Concepts

ETI, White River Jct., VT, 1996.2. N.K. Amineni, A. Engeda, Pressure Recovery in Low Solidity Vaned Dif-

fusers for Centrifugal Compressors, ASME paper, Paper No. 97-GT-472,1997.

3. Y. Senoo, H. Hayami, H. Ueki, Low Solidity Trandem Cascade Diffusers forWide Flow Range Centrifugal Blowers, ASME paper, Paper No. 83-GT-3,1983.

4. C. Xu, R.S. Amano, E. K. Lee, Investigation of an Axial Fan—Blade Stressand Vibration Due to Aerodynamic Pressure Field and Centrifugal Effects,JSME International Journal, Series B, vol. 47, no. 1, pp.75–90, 2004.

5. D.P. Kenny, The History and Future of the Centrifugal Compressor inAviation Gas Turbine, 1st Cliff Garrett Turbomachinery Award Lecture,Society of Automotive Engineers, SAE/SP-804/602, 1984.

6. J.M. Sorokes, Range Versus Efficiency – A Dilemma for Compressor De-signers and Users, IMECE2003-44221, Proceedings of PID Industrial andPipeline Compressors Sessions, ASME - IMECE 2003, November 16–21,Washington D.C. 2003.

7. C. Xu, R.S. Amano,On the Development of Turbine Blade AerodynamicDesign System, ASME IGTI Turbo & Expo, 2001-GT-0443, 2001.

8. C. Xu, R.S. Amano,Turbomachinery Blade Aerodynamic Design and Op-timization, ASME paper, Paper No. GT-2002-30541, 2002.

9. C. Xu, R.S. Amano, A Study of a Single Stage Centrifugal Compressor,Proceedings of ASME POWER: Electric Power Conference, May 2-4,2006, Georgia World Congress Center, Atlanta, Georgia, PWR2006-88023,2006.

10. C. Xu, Impeller for Centrifugal Compressor, United States patent pending,20070059179, 2007.

11. ASME, Performance Test Code on Compressors and Exhausters, PTC 10-97, 1997.

12. A. Hrennikoff, Solution of Problems of Elasticity by the Frame-workMethod, ASME J. Appl. Mech., vol. 8, pp. A619–A715, 1941.

13. BLADEPROTM software, Impact Technologies, LLC, http://www.impact-tek.com, Rochester, NY 14623, 2006.

14. Ansys, Ansys Release 10, Ansys Inc., Pisbbuger, PA, 2005.15. S.W. Mak, M. Botman, Substructures Analysis of Impeller Vibration

Modes, ASME paper, paper NO. 75-DET-112, 1975.16. C. Xu, R.S. Amano, Computational Analysis of Scroll Tongue Shapes to

Compressor Performance by Using Different Turbulence Models, ASMEpaper, Paper No. GT-2007-28224, 2007.

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