See discussions, stats, and author profiles for this publication at: https://www.researchgate.net/publication/245493930
The Development of Centrifugal Compressor Impeller
Article in International Journal for Computational Methods in Engineering Science and Mechanics · June 2009
DOI: 10.1080/15502280903023165
CITATIONS
READS
9
1,024
2 authors:
C. Xu
Ryoichi S. Amano
University of Wisconsin - Milwaukee
University of Wisconsin - Milwaukee
58 PUBLICATIONS 325 CITATIONS
367 PUBLICATIONS 1,352 CITATIONS
SEE PROFILE
Some of the authors of this publication are also working on these related projects:
Internal Cooling of Gas Turbine Blades View project
Energy in water and wastewater treatment View project
All content following this page was uploaded by C. Xu on 12 November 2014.
The user has requested enhancement of the downloaded file.
SEE PROFILE
This article was downloaded by: [University of Connecticut]
On: 28 October 2014, At: 12:12
Publisher: Taylor & Francis
Informa Ltd Registered in England and Wales Registered Number: 1072954 Registered office: Mortimer House,
37-41 Mortimer Street, London W1T 3JH, UK
International Journal for Computational Methods in
Engineering Science and Mechanics
Publication details, including instructions for authors and subscription information:
http://www.tandfonline.com/loi/ucme20
The Development of a Centrifugal Compressor Impeller
a
C. Xu & R. S. Amano
a
a
University of Wisconsin-Milwaukee , Milwaukee, WI, USA
Published online: 11 Jun 2009.
To cite this article: C. Xu & R. S. Amano (2009) The Development of a Centrifugal Compressor Impeller, International Journal
for Computational Methods in Engineering Science and Mechanics, 10:4, 290-301, DOI: 10.1080/15502280903023165
To link to this article: http://dx.doi.org/10.1080/15502280903023165
PLEASE SCROLL DOWN FOR ARTICLE
Taylor & Francis makes every effort to ensure the accuracy of all the information (the “Content”) contained
in the publications on our platform. However, Taylor & Francis, our agents, and our licensors make no
representations or warranties whatsoever as to the accuracy, completeness, or suitability for any purpose of the
Content. Any opinions and views expressed in this publication are the opinions and views of the authors, and
are not the views of or endorsed by Taylor & Francis. The accuracy of the Content should not be relied upon and
should be independently verified with primary sources of information. Taylor and Francis shall not be liable for
any losses, actions, claims, proceedings, demands, costs, expenses, damages, and other liabilities whatsoever
or howsoever caused arising directly or indirectly in connection with, in relation to or arising out of the use of
the Content.
This article may be used for research, teaching, and private study purposes. Any substantial or systematic
reproduction, redistribution, reselling, loan, sub-licensing, systematic supply, or distribution in any
form to anyone is expressly forbidden. Terms & Conditions of access and use can be found at http://
www.tandfonline.com/page/terms-and-conditions
International Journal for Computational Methods in Engineering Science and Mechanics, 10:290–301, 2009
c Taylor & Francis Group, LLC
Copyright ISSN: 1550–2287 print / 1550–2295 online
DOI: 10.1080/15502280903023165
The Development of a Centrifugal Compressor Impeller
C. Xu and R. S. Amano
Downloaded by [University of Connecticut] at 12:12 28 October 2014
University of Wisconsin-Milwaukee, Milwaukee, WI, USA
An impeller is one of the key components of industrial centrifugal compressors and turbochargers. Aerodynamic and structure
designs of the impeller are critical to the success of the whole compressor stages. The requirements for efficiency and operating range
of industrial centrifugal compressors and turbochargers have been
increased dramatically compared with the situation in the past.
The efficiency of a newly developed, low-pressure-ratio centrifugal
compressor has reached the possible level of the machine. However, the efficiency level of an intermediate- and high-pressure
ratio machine still has gaps between the current state-of-the-art
and possible level. The challenge for centrifugal compressor design is to keep the efficiency level at state-of-the-art and increase
the compressor operating range. Increase of the compressor operating range without sacrificing compressor peak efficiency is difficult to achieve. The product globalization requires one product
design, which can be used in all locations. In some counties, due to
the technology differences, electricity frequency variations could
be 10%. Turbocharger compressors work at different rotational
speeds for the majority of the time. The compressor impeller rotating speeds change in certain ranges. The impeller rotating speed
variation makes the impeller structure design more challenging.
In this study, a full-3D impeller was designed to optimize impeller
aerodynamic performance and structure characteristics.
Keywords
Impeller, CFD, FEA
INTRODUCTION
Centrifugal compressors have a wide application in industrial, aerospace, and automobile industries. In a centrifugal compressor the flow enters the compressor axially and then turns in
the radial direction out from the impeller. The flow then enters a
radial annular vaned or vaneless diffuser. The flow exit from the
diffuser needs a volute or collector to deliver the flow to the next
stage or send it to the next components [1–3]. The basic components of the radial compressor are shown in Fig. 1. Unlike an
axial compressor or fan [4, 5], the work input for a centrifugal
compressor is almost independent of the nature of the flow. A
centrifugal compressor can be designed with much higher De
Address of correspondence to R. S. Amano, University of
Wisconsin-Milwaukee, Milwaukee, WI 53201, USA. E-mail: amano@
uwm.edu
Haller number than an axial compressor can achieve. So it is possible for a centrifugal compressor to have a much higher stage
pressure ratio than axial ones. Centrifugal compressors have
wide applications for a mass flow rate less than 10 kg/s [1, 5].
The small flow centrifugal compressors have wide applications in turboshaft aircraft engines, turbochargers, and industry.
However, being in a low-power class, these compressors need
to be inexpensive to manufacture and operate, requiring that the
compressor be a simple design having fewer parts and relatively
large tolerance. Moreover, the five-axis machine is now a common tool for impeller machining, but most other parts should
be fabricated by using other types of machining to reduce the
manufacturing costs.
The turbomachinery industry is increasingly interested in using optimization procedures that enable enhanced compressor
efficiency and widen operating ranges. During the compressor development, designers always compromise between a peak
attainable efficiency and an overall operating range [6]. Compressor design normally starts with a meanline program at each
individual operating point on a map, then throughflow calculation is performed, and finally the impeller, diffuser, and volute
are designed. In this study, a recently developed turbomachinery
viscous optimal method [7, 8] for axial machines was further
extended to a centrifugal compressor design [9]. The main focus
of this study lies in the development of a small flow compressor
where flow coefficient φ = Q/N/D3 = 0.145. Q, N, and D are
the volume metric flow (m3 /s), rotational speed (rps), and impeller tip diameter (m). The design pressure ratio and the flow
rate are 3.65 and 0.75 kg/s, respectively, at the design condition.
The total to static efficiency required to be higher than 84% and
the stability operating range (SB) great than 38%. The compressor design employs a global optimization viscous process
for achieving efficiency and stability targets [8]. Good surge
margins were achieved without use of a variable geometry for a
stable operation. Special attention has been paid to the tip clearance profile to permit large clearance without too much lost
efficiency during the impeller design. The compressor developed in this study consists of three major parts: an impeller, low
solidity diffuser, and volute. In this study, particular attention
was paid mainly to impeller design and analyses.
290
291
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
TABLE 1
Main design information of the test compressor
Shaft speed N = 60000 rpm
Impeller outlet radius R2 =
82.90 mm
Out blade angle β2 = 55◦
Downloaded by [University of Connecticut] at 12:12 28 October 2014
Impeller blade number Zi =
19
Diffuser inlet radius R3 =
88.70 mm
FIG. 1. Centrifugal compressor.
A variable speed centrifugal compressor impeller with a compound lean blade was developed [9, 10]. The newly developed
impeller defines a blade pressure side that extends from the
leading edge to the trailing edge and is convex near the leading
edge and flat near the trailing edge. The impeller was leaned opposite to rotation at the inlet portion and leaned back at exit. The
clearance between impeller and casing is non-uniform, defined
by a high-order equation. The flow and structure analyses were
performed. Some basic performance tests and structure modal
tests were performed. The analyses agreed with test results.
EXPERIMENTAL STUDY
Two impeller designs were installed with the same diffuser
and scroll stage. The experimental studies were performed in
the Centrifugal Compressor Development Laboratory. The performer test and data acquisition procedures were based on the
ASME performance test procedure [11]. The test compressor
consists of: inlet pipe, impeller, vaneless diffuser, low solidity
vaned 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 the
discharge pipe of the compressor. The main dimensions of the
test compressor are listed in Table 1. Reference [9] provides
more 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 were
kept under 2.2% and 2.5%, respectively, with the same confidence level based on the system uncertainty analyses [9]. The
performance characteristic was tested through five total temperature and total pressure measurements at the inlet of compressor
and discharge of the volute exit cone for both impellers. Due to
the motor system problems after running the traditional impeller
performance test, the off-design point test for a new full-threedimensional impeller was not able to performance. Only design
Mass flow rate m
= 0.77 kg/s
Impeller outlet width b2 =
7.85 mm
Design pressure ratio π =
3.60
Numbers of diffuser vanes
Zd = 9
Volute inlet radius R5 =
170.00 mm
point performance was tested for a full-three-dimensional impeller. More testing for off-design point performance will be
conducted after fixing the motor problem. For determining the
static pressure ratio, five pressure transducer taps were mounted
on the circumferentially distributed wall at discharge of volute
exit cone, which were used to measure the pressure. The mass
flow was measured by using an ASME [10] nozzle located at
the end of the discharge pipe. The tip clearance of the impeller
at tip was measured with an alumina pin. The design tip running clearances at inlet and exit are about 0.2 mm and 0.1 mm,
respectively. The performance test results are discussed in the
CFD calculation section.
For comparing the structure analyses with the experimental
results, vibration tests were performed for the full 3D impeller
utilizing a LDV (Laser Doppler Velocitmeter) and an impact
hammer. A calibrated impact hammer was used to excite the impeller at different points and the resulting frequency and mode
shape were measured by fast Fourier transform (FFT) and circular sampling of the part. Post processing yields a spectral
response pattern and mode shape. The frequency test results are
reported in the following section.
FIG. 2. Test rig of a single stage compressor.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
292
C. XU AND R. S. AMANO
FIG. 3. FEA boundary conditions and meshed for different impeller blades.
FEA AND MODAL ANALYSES
In a modern centrifugal compressor, due to increased requirements of compressor performance, the blade thickness has
a trend to reduce. This can be achieved by the selection of better
material and conducting more detailed finite-element analysis
(FEA).
Finite element analysis is a powerful tool to help find correction factors that make more cost-effective, simplified analyses
as accurate as possible [12]. Stress analyses of the centrifugal
impeller wheel during the design stage helps diagnose possible
design problems and avoid failure. The finite element method
can identify high-stress locations as well as vibrations in the
centrifugal compressor impellers. Integration optimization between FEA and CFD is more popular in industry for turbomachinery designs [7,8]. Optimizations guided design changes that
improved the integrity of the compressor.
In this study, a preprocess BLADEPROTM [13] was used for
Ansys [14] solver. The preprocess allows the user to specify Radial (R), Tangential (T), and Axial (A) boundary conditions for
cover-face nodes for both static (single-blade steady stress) and
dynamic (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 condition, Radial constrain boundary condition and Radial and Axial constrain boundary conditions. The temperature and pressure
profiles on the blade were calculated based on the CFD analyses.
After generating the mesh and setting boundary conditions, the
stress and modal analyses were performed. The material properties such as Youngs Modulus, material density and Poisson’s
ratio and thermal properties such as coefficient of thermal expansion were selected from BLADEPROTM material database
(Ti-6A1-4v Titanium Alloy). Calculations were performed on a
Dell Precision 490 personal computer. The calculation time for
stress analysis was about 30 seconds.
The calculation showed that overall von Mises stress is similar for both full three-dimensional blade and traditional threedimensional design, those of which are the cases without bow
and compound lean features. The von Mises stress contours near
the 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 3D
blade and is about 82.3 ksi for traditional blade. It is shown that
the bore stress does not change too much for both designs. The
von Mises stress contours near the root of the trailing edge is
shown in Fig. 5. As shown in Fig. 5, the maximum von Mises
stress for the full 3D blade is a little higher than the cases for
traditional blades. This is because the full 3D blade trailing edge
stiffness may be smaller than a traditional blade. Less stiffness
will cause more deflection and increase the stress. However, all
the von Mises stress level is well below the yield strength of
the material. The maximum von Mises stress for full 3D blade
peak stress near the leading edge is about 64ksi and is about
52ksi 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.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
293
FIG. 5. von Mises stress contour near trailing edge for different impeller blade designs.
stress mainly depends on how much material near the leading
edge area is used. The full 3D blade design has more material
near the leading edge area due to the compound lean and bow.
Additional material increases the centrifugal force during the
rotation and causes higher stress. However, the maximum stress
near the root of the leading edge for both designs is far below yield strength. It is shown that both designs meet the stress
requirements.
After stress analyses were completed, an at-speed modal
analysis was performed in which a single-blade steady stress
analysis was first performed to include the stress stiffening effects in frequency prediction. Stress stiffening effects from a
single-blade stress analyses were carried over into the frequency
calculation. At-speed modal analysis performs frequency prediction at a given shaft speed where both the “stress stiffening”
and “spin softening” effects are included in frequency
prediction.
A cyclic sector can be modeled using two options: the reduced order model (ROM) and the full cyclic model. ROM
uses super-elements for frequency prediction whereas the full
cyclic model uses all the elements in the cyclic sector for frequency prediction. The full wheel method is used when there
is a non-cyclic structure; for example, a wheel with unequal
group lengths. The full cyclic model uses all the elements in the
cyclic sector for frequency prediction. In this study, an at-speed
full cyclic modal analysis was performed on a Dell Precision
490 personal computer after stress analyses. Modal analyses for
each case took about 2500 seconds.
The interference diagram is a plot of Frequency vs. Nodal
Diameter that is used to identify potential resonance situations
FIG. 6. von Mises stress contour near leading edge root for different impeller blade designs.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
294
C. XU AND R. S. AMANO
FIG. 7. Interference diagram for different impeller blade designs.
[15]. Each mode is represented by a dot on the interference
diagram. The impulse lines in the interference diagram were
constant speed lines corresponding to high and low limits of
the running speeds of the machine. Intersection of this constant
speed line with one or more dots indicates potential resonance
conditions. The modal analyses results are plotted in the form of
interference 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 profiles and thickness; however, it successfully meets the frequency
margin requirements.
The frequency prediction accuracy is dependent on the mode
shape and level of the frequency; simple modes like 1st bending
FIG. 8. Sketch of the computational mesh.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
295
FIG. 9. Convergent history of RMS mass.
and torsion can be predicted more accurately than higher complex modes. The full three-dimensional blade mode calculations
are more difficult than traditional three-dimensional blades. For
validating the calculations, the modal testing for a full 3D wheel
was conducted. It can be seen that the calculation results agree
with the measurements.
FIG. 10. Comparison of the performance curves.
CFD CALCULATIONS
The single stage compressor, from inlet pipe, impeller, vaned
diffuser, vanless diffuser and volute, is computationally analyzed using a Navier-Stokes solver on a fully 3-D viscous turbulent flow solver CFX10.0. The geometrical discretization of
the compressor stage is made for the computational analysis.
FIG. 11.
Tip relative Mach number distribution near shroud.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
296
C. XU AND R. S. AMANO
FIG. 12. Relative Mach number contour near suction side of the blade.
FIG. 13. Relative Mach number contour near pressure side of the blade.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
FIG. 14. Relative Mach number contour near mid plan between the blades.
FIG. 15. Static pressure contour near suction side.
297
Downloaded by [University of Connecticut] at 12:12 28 October 2014
298
C. XU AND R. S. AMANO
FIG. 16. Static pressure contour near pressure surface.
FIG. 17. Radial velocity contour near suction surface.
Downloaded by [University of Connecticut] at 12:12 28 October 2014
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
299
FIG. 18. Radial velocity contour near pressure surface.
Structure hexahedral cells are generated to define all parts of
the compressor as shown in Fig. 8. The meshed size for inlet
pipe, impeller, diffuser, and volute are 29,940, 597,902, 497,314
and 768,543 nodes, respectively. The interface surface between
inlet-impeller and impeller-diffuser is modeled by using the
frozen rotor method, i.e. the relative orientation of the two interface components across the interface is fixed. The calculation
of the downstream surface of the interface plan was based on
the average mixing plane approach. The calculation assumed to
be converged when RMS of flow, pressure, and turbulence parameters smaller than the 7th order of original errors as shown
in Fig. 9.
The turbulence model is one of the factors to determine the
success of the CFD analyses. Different turbulence models were
tested for single stage centrifugal compressor flow calculations
[16]. It is shown that both k-ε two equations and zero equation
turbulence models provided reasonably good results, as shown
in Fig. 10. The turbulence model used in this study is based on a
zero equation model; only an algebraic equation was used to calculate the viscous contribution from turbulent eddies. The inlet
boundary conditions were enforced by using the total pressure
and total temperature. The flow entering the pipe is assumed to
be normal to the inlet surface. The outlet boundary condition
was applied by using a variable static pressure proportional to
the kinetic energy at the outlet. The flow rate of the compressor
is changed by modifying the static pressure to kinetic energy
ratio at the outlet condition, which simulates different closing
positions of the butterfly valve employed in the tests. In the
calculation, no-slip boundary conditions were imposed over the
impeller blades, diffuser airfoils, and all solid walls.
The calculations were performed in an HP 8000c workstation. 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 conditions. The calculation was assumed to be convergent when the
ratio between the sum of the residuals and the sum of the fluxes
for a given variable in all the cells is reduced to at least six orders
of magnitude. Intensive grid size dependence tests were carried
out [9] and overall compressor performance was compared with
different mesh sizes. The final mesh size was set when the mesh
size increased; the mass flow rate changed less than 0.5%.
Two types of turbulence models, standard k-ε two equation
and zero equation turbulence models, were used to calculate the
compressor performance for traditional impeller. The results
are plotted in Fig. 10. It is shown that the compressor developed by this research has a wide operating range of 40% with
high efficiency in all operating conditions. It is shown that both
turbulence models provided good predications for compressor
efficiencies and head coefficients compared with the test, except
the compressor operated near choke. For the compressor operated near choke, computations overpredicted the performance
compared with tests. For full 3D design, the calculation shows
a little performance advantage, better efficiency and head at the
same flow conditions, compared with traditional impeller design. However, the tests near design flow show that there are
no significant benefits for both efficiency and head coefficients
near design point. The test showed a little benefit near choke and
surge operations. To reduce the test tolerances, a few repeated
tests were performed and similar results were obtained. Upon
further inspecting the impellers, it was found that the impeller
surface finish is about 6.3 micrometers near hub and about 3.2
300
C. XU AND R. S. AMANO
Downloaded by [University of Connecticut] at 12:12 28 October 2014
FIG. 19.
Relative flow angle contour at blade exit.
micrometers on the blades. Because the calculations assumed
that all walls on the impeller are a smooth wall, the friction
loss estimations may be smaller than the test impeller. When a
compressor is operated at a near design condition, both full 3D
impeller and traditional impeller have no massive separation;
the friction loss for the full 3D impeller is larger than the traditional blade due to bowed and lean features increasing the wet
area. Further examination of the impeller calculations at design
point for the compounded lean 3D impeller and conventional
impeller showed that the total to total efficiency for 3D impeller
and traditional impeller were 93.33% and 92.99%, respectively.
It can be seen that the gain of the efficiency for 3D design is
about 1/3%. This performance gain in the impeller becomes less
significant for a whole stage. If the test part surface finish is not
very high, the effect of the performance due to a 3D design will
be washed out by frictional loss.
The tip relative Mach number contours near the tip of the
impellers for both full 3D and traditional designed impeller
are shown in Fig. 11. It is shown that the traditional design
leading edge area had a relatively larger Mach number compared with the full 3D design impeller. This may indicate that
the full 3D design had a smaller portion of the flow passing through the blade tip range. It is favorable for impeller
FIG. 20.
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 better
efficiency.
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 a
relatively larger high Mach number area than the conventional
blade near the suction side of the blade in the middle of the flow
path. At a near inlet area the full 3D design depicts a smaller
high Mach number area than the traditional impeller design.
On the pressure side, the low Mach number region of the full
3D impeller is smaller than the conventional design. Mid-span
Mach number distributions also show that a full 3D design provides more uniform Mach number distributions. The uniform
Mach number distributions would benefit the compressor performance.
The static pressure contours near the pressure and suction
side are shown in Figs. 15 and 16. The static pressure contours
show that the traditional blade has less static pressure rise compared with a lean blade. Figures 17 and 18 show the relative
radial velocity contour near suction and the pressure surface of
the impeller blade, respectively. It is shown that there is a larger
high radial velocity region for the traditional impeller both near
Secondary velocity contour at blade exit.
THE DEVELOPMENT OF A CENTRIFUGAL COMPRESSOR IMPELLER
301
Downloaded by [University of Connecticut] at 12:12 28 October 2014
FIG. 21. Relative Mach number contour at blade exit.
the suction and pressure side at about 50% of the flow path location of the impeller. It also shows that this higher velocity region
is also closer to impeller shroud for traditional blade impeller.
It is interesting to point out that the radial velocities near the
shroud at 50% flow path location drop very fast near the impeller
exit. At the impeller exit, the radial velocities near shroud are
very low.
The flow angle contour at blade exit plan in Fig. 19 shows
that the compounded lean full 3D blades design provides a
more uniform exit angle distribution. The larger angle flow area
is smaller for the full 3D design. This characteristic will benefit
the vaned diffuser performance. This allows a 3D design to get
a little more efficiency compared with a traditional design. The
secondary flow contour at the exit of the impeller is shown in
Fig. 20. It is shown here that both designs give a single vortex
structure and the main secondary flow center is located near the
suction side of the blade. In the traditional design, the vortex
center is located closer to the shroud side and also shows much
stronger secondary flows. The relative Mach number distribution in Fig. 21 shows that the regions for both high and low Mach
number areas are larger for the traditional design. It indicated
that the flow is less uniform at the impeller exit for traditional
design.
CONCLUSIONS
In this study, a new full 3D centrifugal impeller was designed
and analyzed for both structure and flow. Calculations showed
that a full 3D blade is both advantageous in a structure vibration
and an aerodynamic performance. The test results show that
the full 3D impeller did not show any significant benefits in
performance near the design point due to the increase in friction
losses on the surface of the bowed blades, whose surface area
is greater than that of the conventional impeller. In cases of
near surge and choke, full 3D impeller separation is reduced
in the bowed blade impeller so that the stagnation pressure
ratio and efficiency are increased. Structure analyses showed
that the bowed blade could change the impeller into higher
order modes separations. Natural frequencies of the impeller had
View publication stats
different values compared with the traditional blade. This feature
offers benefits that previously have not been reported in the
literature.
REFERENCES
1. 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 Diffusers for Centrifugal Compressors, ASME paper, Paper No. 97-GT-472,
1997.
3. Y. Senoo, H. Hayami, H. Ueki, Low Solidity Trandem Cascade Diffusers for
Wide 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 Stress
and 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 in
Aviation 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 Designers and Users, IMECE2003-44221, Proceedings of PID Industrial and
Pipeline Compressors Sessions, ASME - IMECE 2003, November 16–21,
Washington D.C. 2003.
7. C. Xu, R.S. Amano,On the Development of Turbine Blade Aerodynamic
Design System, ASME IGTI Turbo & Expo, 2001-GT-0443, 2001.
8. C. Xu, R.S. Amano,Turbomachinery Blade Aerodynamic Design and Optimization, 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 1097, 1997.
12. A. Hrennikoff, Solution of Problems of Elasticity by the Frame-work
Method, ASME J. Appl. Mech., vol. 8, pp. A619–A715, 1941.
13. BLADEPROTM software, Impact Technologies, LLC, http://www.impacttek.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, ASME
paper, Paper No. GT-2007-28224, 2007.