-
JongSik OhSenior Aerodynamic Design Engineer
e-mail: [email protected]
Charles W. BuckleyEngineering Manager
e-mail: [email protected]
Giri L. AgrawalPresident
e-mail: [email protected]
R&D Dynamics Corporation,
49 West Dudley Road,
Bloomfield, CT 06002
Numerical Investigationof Low Solidity Vaned DiffuserPerformance
in a High-PressureCentrifugal Compressor—Part III:Tandem VanesAs
Part III, following the authors’ previous studies, the aerodynamic
performance of twodifferent tandem LSDs (low solidity diffusers),
Tandem (A) and (B), in a high-pressurecentrifugal compressor was
numerically investigated over flow rates from impeller choketo
minimal flows available in computation. Tandem (A) was of
conventional design wherethe first row came directly from the
authors’ previous studies (Part I Oh and Agrawal,2007, “Numerical
Investigation of Low Solidity Vaned Diffuser Performance in a
High-Pressure Centrifugal Compressor - Part I : Influence of Vane
Solidity,” ASME Paper No.GT2007-27260, and Part II: Oh et al.,
2008, “Numerical Investigation of Low SolidityVaned Diffuser
Performance in a High-Pressure Centrifugal Compressor - Part II :
Influ-ence of Vane Stagger,” ASME Paper No. GT2008-50178) selected
as the highest effi-ciency vane at design flow, and the second row
was designed to be added consideringflow conditions at the exit of
the first row vane. Tandem (B) followed a creative patent-pending
concept where the number of the first row vanes was doubled with
much smallervane chord keeping a low solidity. A position parameter
of RCP (relative circumferentialposition) was introduced to see the
effect of the relative location of the second row vane.Using an
in-house Navier-Stokes solver with finite volume time marching
methods, over-all performance was predicted to be compared with
each other. Detailed investigation onthe behavior of the static
pressure recovery and the total pressure loss coefficient in
bothdiffuser designs helps determine why Tandem (A) design is
better and the case ofRCP¼ 0.3 gives the best performance. [DOI:
10.1115/1.4006300]
Introduction
As the LSD in centrifugal or mixed-flow compressors hasbecome
popular, especially in industrial applications, the aerody-namic
designer has to determine important design parameters,such as the
solidity, the vane profile, the vane stagger, and the ra-dial
position of the vane. Many experimental and computationalresearch
studies have been conducted on this area, but not manyparametric
studies have been tried in a systematic way which arereally needed
for designers. The high cost of experimental studieswould be one
reason for the lack of such a parametric research,but the CFD
(computational fluid dynamics) approach providesreasonably accurate
predictions in a cost-effective way as long asit is limited to
finding out the trend of overall compressor aerody-namic
performance. The authors have studied numerically theinfluence of
those design parameters for a single row LSD with anidentical
centrifugal impeller through studies in series [1,2]. Thepresent
study, as Part III, is about the design parameters for a tan-dem
LSD.
The tandem vane is another option in the LSD design toincrease
the static pressure recovery more than a single row caseby adding
the second row of vanes. Pampreen [3] made a compar-ison between
three-row vanes and a single channel-wedge dif-fuser, and argued
that the tandem vane was superior in testperformance. A wider
operating range and higher efficiency wasfound for the tandem LSD
than for the channel diffuser. Senoo [4]
tested a tandem LSD with a blower impeller and found only asmall
gain in performance over a single row LSD, as summarizedby Osborne
and Sorokes [5]. Because the inlet flow condition ofthe second row
vane is not the same as that of a single row case, adifferent set
of vane profiles and vane staggers should be selectedto diffuse
with minimum losses. A critical design parameter forthe performance
of the tandem vane is the relative circumferentialposition (RCP) of
the second row vanes (See Fig. 1) because testresults have shown
that there is a definite preference for position-ing the second row
relative to the first. Very limited systematicdesign information is
available for tandem LSDs in publishedreferences. Japikse [6]
quotes unpublished test data by Pampreenthat lower loss was
achieved when the suction surface of the
Fig. 1 Definition of RCP
Contributed by the International Gas Turbine Institute (IGTI) of
ASME for publi-cation in the JOURNAL OF TURBOMACHINERY. Manuscript
received July 13, 2011; finalmanuscript received July 26, 2011;
published online September 4, 2012. Editor:David Wisler.
Journal of Turbomachinery NOVEMBER 2012, Vol. 134 /
061025-1Copyright VC 2012 by ASME
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
second row leading-edge was close to the pressure
surfacetrailing-edge of the first row (when RCP is close to 1.0).
Seleznevand Galerkin [7] obtained the best aerodynamic loading with
thesecond row placed 10% from either side of the trailing-edge of
the
first row, that is RCP¼ 0.1 or 0.9, from analytical results.
Maxi-mum efficiency was found when RCP¼ 0.1, and the next best
effi-ciency at RCP¼ 0.9, which differs from the above
statement.Bandukwalla [8] proposed an interesting design concept of
thetandem LSD where a high number of vanes, between 19 and 22,was
assigned to the first row, with a split tandem vane for the sec-ond
row. The first row had to have a much smaller chord to keepthe
solidity less than 1.0, and the vane number of the second rowwas
decreased to half the number of the first row. The conceptintended
to provide the combined advantages of the low solidityand high
solidity diffuser because it was believed that, for goodefficiency,
the diffuser vane number should be 10% to 50% morethan the impeller
blade number.
Among many design parameters to select in the tandem LSD,the
authors need to know what range of RCP gives high perform-ance.
Furthermore, to find out the feasibility of Bandukwalla’sconcept
[8], another design version (Tandem (B)) was added tothe
conventional original design (Tandem (A)), as shown inFig. 2. As
RCP was varied from 0.0 to 0.9 for six cases, overallcompressor
performance was numerically investigated for bothdesign versions of
Tandem (A) and (B) at design speed of rotationfrom impeller choke
to minimal flows available in computation.
Centrifugal Compressor
The centrifugal compressor in this study is, as shown in Figs.
2and 3, from a marine use turbocharger for medium-class shipengines
whose design pressure ratio (total-to-static) is 4.0 anddesign
isentropic efficiency (total-to-static) is 80%. The design airmass
flow rate is 3.0 kg/s at design speed of 34,000 rpm. The
Fig. 2 Front view of Tandem (A) and (B) when RCP 5 0.5
Fig. 3 Centrifugal compressor geometry in meridional view
061025-2 / Vol. 134, NOVEMBER 2012 Transactions of the ASME
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
unshrouded open impeller has 18 full blades, and is 45-deg
back-swept. Tandem (A) is of a conventional original design where
thefirst row comes directly from the authors’ previous studies
(Part Iand Part II), selected as the highest efficiency vane at
design flow,and the second row is designed to be added considering
flow con-ditions at the exit of the first row vane. Tandem (B), as
mentionedbefore, follows the creative concept of Bandukwalla [8]
where thenumber of the first row vanes is doubled with much smaller
vanechord keeping a low solidity. Detailed information of both
tandemvanes is shown Table 1. The second row of Tandem (A) has
rela-tively lower solidity because the fixed location of the volute
inletlimits radial space required for the second row in the present
case.In other words, if the volute inlet location is flexible,
moreimproved performance may be possible. Because the stagger
inTable 1 is an angle obtained using a straight line in the
radialplane, the first rows of Tandem (A) and (B) have different
stag-gers despite the same design angle of attack. The selected
LSDvane profiles are NACA65-(4A10)06 for the first row,
andNACA65-(12A10)10 for the second row, as one of the most popu-lar
airfoil combinations, which are transformed onto a radialplane.
Original thickness distributions of both vanes were how-ever
intentionally increased to provide a way of cost-effective
fab-rication. A vaneless space downstream of the impeller is
gentlycontracted to give a pinch for improving flow stability. A
muchstronger contraction is intentionally provided at CFD exit
bound-ary, only for calculation purposes, which is usually required
toapproach lower flows by partly eliminating reverse flows.
Numerical Method
Compressible flow in a whole domain from CFD inlet boundaryto
CFD exit boundary, shown in Fig. 3, was analyzed using anin-house
code, CNSTURBO [9,10], that employs the finite vol-ume method with
4-step Runge-Kutta time integration schemeand the 2nd/4th-order
artificial dissipation damping. It has beenextended to cover a
cut-off trailing-edge of blades and a realisticrectangular tip
clearance region using multi-block grid capability,and to add the
k-omega equation model, used in the present studyas a turbulence
closure. Due to its original features of time march-ing methods,
upstream boundary total pressure and temperatureare given with flow
directions, and static pressure is imposed asthe exit boundary
condition to obtain a converged mass flow rateas part of solution.
In grid generation, normally about 206,000nodes and about 390,000
nodes were used to build the impellerand tandem LSD domains,
respectively, using the H-type struc-tured grids, as shown in Fig.
4. A grid sensitivity study had beenmade in Part I [1] where
doubling the sizes of the computationalgrids had produced a
difference in performance within 1.6%range, and of course much more
computation time and memoryrequirement. The current grid sizes are
therefore recognized to bereasonable and efficient because the
authors are only interested ina steady state solution for overall
compressor performance tobuild a supporting design guide. Both
impeller and diffuserdomains were combined to produce a single
domain with the so-called stage interaction (or mixing plane)
scheme applied whereall computed flow properties were
circumferentially averaged,while the spanwise variation was still
preserved, for a steady state
solution in a simple way at the rotational/stationary
interfacelocated at halfway distance. A 5% of span was consistently
treatedas running tip clearance from impeller inlet to exit. By
varyingstatic pressures at the exit boundary, computational flow
pointswere shifted from choke toward stall. In the present study,
thelowest mass flow point for each configuration does not mean
atrue stall/surge location, because any reverse flows occurring
forlower flow rates in the numerical computation become an
obstacleto solution convergence. It has to be understood that each
lowestflow in the present study is a minimum flow with an
acceptabletolerance of solution convergence. Steady numerical
solution atany flow less than each lowest flow was not converged
success-fully. The convergence criteria used in this study is that
the solu-tion was regarded as converged when the normalized
residual, ameasure of local imbalance of each conservative control
volume,fell below 1.0� 10�5. In data reduction, all performance
parame-ters were evaluated using mass-averaged pressure,
temperatureand velocities at any plane sections.
Results and Discussion
Figures 5 and 6 show the total-to-static pressure ratio and
isen-tropic efficiency distributions, respectively, when RCP
variesfrom 0.0 to 0.9 for both Tandem (A) and (B) cases. The
pressureratio was calculated from the impeller upstream up to the
locationof the volute inlet. Irrespective of RCP, Tandem (A)
providesmore elevated pressure rise and efficiency than Tandem (B).
InTandem (A), the case of RCP¼ 0.3 shows the highest pressurerise
with a wide operating range. As RCP moves from 0.3 to 0.9crossing
over 0.0 (that is, as the second row vane moves to thecounter-clock
wise direction from RCP¼ 0.3 in Fig. 1), the per-formance drops
accordingly. However, the case of RCP¼ 0.7 hasthe lowest pressure
rise despite the widest range of operation, andinterestingly it is
inferior to the case of RCP¼ 0.5. In Tandem(B), the operation range
is severely restricted in the case ofRCP¼ 0.1. The cases of RCP¼
0.7 or 0.9 provide the best overallperformance unlike those in
Tandem (A).
Figure 7 was produced at design flow in both Tandem (A) and(B)
to make a comparison among the pressure ratio, the
isentropicefficiency and the numerical operation range. The
numerical oper-ating range is defined as the ratio of mass flow
rate changebetween maximum and minimum flow rates to maximum
flowrate. In Tandem (A), considering all three performance
Fig. 4 Computational grids for whole domain when RCP 5 0.5
Table 1 Diffuser vane information
Tandem (A) Tandem (B)
1st-row 2nd-row 1st-row 2nd-row
Number of vanes 11 11 22 11Solidity 0.71 0.591 0.613 0.72Stagger
(deg) 19.58 28.44 24.75 25.36routlet/rinlet 1.2 1.2 1.08 1.2rvolute
inlet/routlet 1.08 1.2
Journal of Turbomachinery NOVEMBER 2012, Vol. 134 / 061025-3
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
parameters, the case of RCP¼ 0.3 shows the best performance,and
the next best performance is found at the case of RCP¼ 0.9.The
cases of RCP¼ 0.1 and 0.0 show as good pressure rise and
ef-ficiency at design flow as that of RCP¼ 0.9, but has poor
operat-ing ranges. The largest range of operation is found at the
case ofRCP¼ 0.7, but the case shows the worst performance at
designflow. In Tandem (B), all the cases provide lower design
perform-ance than Tandem (A) as mentioned earlier. It is
interesting tonote that the case of RCP¼ 0.0 shows a good operating
range de-spite its lower pressure rise and efficiency.
To see more details at each row of both tandem LSDs, Figs. 8and
9 show the distributions of static pressure recovery factor(CP) in
the first and the second row, respectively. Figures 10 and11 are
the distributions of total pressure loss coefficient (LC) inthe
first and the second row, respectively. The total pressure
losscoefficient is defined as a ratio of total pressure drop to
upstreamdynamic pressure, and the static pressure recovery factor
isdefined as a ratio of static pressure rise to upstream
dynamicpressure.
All cases in Tandem (B) show much lower static pressure
re-covery in the first row because of smaller vane chord (Fig. 8),
and
they are high in total pressure loss coefficient in the second
row(Fig. 11), which drives Tandem (B) away from acceptable
per-formance. In static pressure recovery in the second row (Fig.
9)and total pressure loss coefficient in the first row (Fig. 10),
no re-markable difference is found between Tandem (A) and (B).
Fig. 6 Compressor isentropic efficiency characteristic
Fig. 8 Static pressure recovery characteristic in the first
row
Fig. 5 Compressor pressure ratio characteristic
Fig. 7 Compressor performance at design flow
061025-4 / Vol. 134, NOVEMBER 2012 Transactions of the ASME
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
In Tandem (A), the case of RCP¼ 0.9 shows the highest
staticpressure recovery in the first row, while the case of RCP¼
0.1 hasthe highest in the second row. Some limited references have
statedthat either case would be considered optimal in aerodynamic
per-formance, and recommended in design.
As shown in Figs. 8 and 9, however, either case of RCP¼ 0.1or
0.9 fails to provide reasonably high pressure recovery in bothfirst
and second rows, on account of the absence of uniform load-ings at
each vane row. Moreover, the case of RCP¼ 0.1 shows thehighest
total pressure loss in the first row, according to Fig. 10.When
RCP¼ 0.7 or 0.9, in other words, when the second rowleading-edge is
approaching the pressure surface of the first row,the static
pressure recovery is rising in the first row, but fallingdown in
the second row. The case of RCP¼ 0.3, which providesthe best
overall performance already found in Figs. 5 and 6, hasthe lowest
total pressure loss in the first row, and has good levelsof static
pressure recovery around 0.6 in both rows.
Static pressure contours at mid-span at design flow rate
areshown in Figs. 12(a) and 12(b), respectively for Tandem (A)and
(B), where the static pressure contour values are normalizedby
compressor upstream total pressure. At first sight, Tandem(B) fails
to provide as much static pressure recovery, comparedto Tandem (A),
because of the much smaller chord of the firstrow vane of low
solidity. The worst non-uniform distribution ofblade loadings in
Tandem (B) is found at RCP¼ 0.1 which con-tributes to the
restriction of the machine operation range, asalready seen in Figs.
5 and 6. As RCP increases, the blade load-ings become more uniform,
and the case of RCP¼ 0.7 or 0.9 inTandem (B) provides the best
loading distribution around thevanes.
In Tandem (A), it is clear that the presence of the
leading-edgeof the second row vane generates a local jump in static
pressuredistributions due to the formation of stagnation flow. As
RCPincreases from 0.0 to 0.1, the static pressure jump expands to
therear portion on the suction surface of the first row vane,
resultingin a sudden rise of static pressure on the suction surface
of the firstrow vane and a sudden drop on the suction surface of
the second.At RCP¼ 0.3, however, the non-uniform blade loading is
muchweakened because the second row vane is located far awayenough
to damp the static pressure jump. However, fromRCP¼ 0.5 to 0.7, the
second row vane destroys again the staticpressure rise on the
pressure surface of the first row vane, becauseof accelerating flow
in the reduced passage. Especially whenRCP¼ 0.5, even though the
second row vane is positioned exactlyhalfway to the first row
channel, its presence accelerates the flowaround the pressure
surface of the first row vane resulting in poorstatic pressure
recovery of the first row. When RCP¼ 0.9, how-ever, the flow
acceleration is much weakened by a small gapbetween the two vanes
resulting in the best static pressure recov-ery of the first row,
but the static pressure recovery of the secondrow drops.
The authors were also interested in performance comparisonof
LSDs with channel-wedge diffusers which are well known asone of
high efficiency diffusers. For the same impeller, twodifferent
channel-wedge diffusers were designed for the perform-ance
comparison, and the results were summarized in the Appen-dix. The
highest efficiency was confirmed with the channel-wedge diffuser,
but the present Tandem (A) was found moreattractive for both design
issues of the efficiency and the operat-ing range.
Fig. 10 Total pressure loss characteristic in the first row
Fig. 11 Total pressure loss characteristic in the second rowFig.
9 Static pressure recovery characteristic in the second row
Journal of Turbomachinery NOVEMBER 2012, Vol. 134 / 061025-5
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
Conclusions
As Part III, the aerodynamic performance of two different
tandemLSDs in a high-pressure centrifugal compressor was
numericallyinvestigated. When the relative location of the second
row vane wasvaried using a position parameter of RCP from 0.0 to
0.9 in Tandem(A) and (B), the followings are drawn as concluding
remarks.
(a) Irrespective of RCP, Tandem (A) provides more
elevatedpressure rise and efficiency than Tandem (B), because
Tan-dem (B) has much lower static pressure recovery in the firstrow
due to smaller vane chord, and higher total pressureloss
coefficient in the second row.
(b) Considering all three design parameters of pressure
ratio,efficiency and operating range in Tandem (A), the case ofRCP¼
0.3 shows the best performance, and the next ac-ceptable
performance is found in the case of RCP¼ 0.9.
(c) The case of RCP¼ 0.9 in Tandem (A) fails to provide
rea-sonably high pressure recovery in the second row, onaccount of
the absence of uniform blade loadings.
(d) The case of RCP¼ 0.1 in Tandem (A) shows a reasonablelevel
of pressure recovery in both vane rows, comparable tothe case of
RCP¼ 0.3, but has a limited operating range.
NomenclatureSolidity ¼ (vane chord)/(tangential pitch)
LC ¼ total pressure loss coefficient (¼ a ratio of totalpressure
drop to upstream dynamic pressure)
CP ¼ static pressure recovery factor (¼ a ratio of
staticpressure rise to upstream dynamic pressure)
23 ¼ between the impeller exit and the first-row diffuservane
inlet
34 ¼ between the first-row diffuser vane inlet and exit45 ¼
between the second-row diffuser vane inlet and exit56 ¼ between the
second-row diffuser vane exit and the
volute inlet(See Fig. 3)
Appendix: Comparison With Channel-Wedge Diffuser
Through a series of studies from Part I to Part III, a
parametricinvestigation on some important design variables in the
LSD foran identical high-pressure centrifugal impeller has been
success-fully completed using CFD work. Another interest that
attractedthe authors was about performance gap from channel-wedge
dif-fusers which are well known for higher efficiency. At first,
anoptimal channel was designed for the given geometry which isnamed
“Channel-wedge Optimal,” as shown in Table 2 but vanesare choked at
far less than design flow rate despite higher per-formance. In
order to open throat area more, the second channelwas designed with
fewer vanes and changes of stagger which isnamed as
“Channel-wedge,” and has some choke margin at designflow rate. Both
channel-wedge diffusers were calculated with the
Fig. 12 Static pressure contours at midspan at design flow
061025-6 / Vol. 134, NOVEMBER 2012 Transactions of the ASME
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
same impeller using the same CFD code from deep choke to
thesmallest flow that the calculation allows at design speed by
speci-fying an exit boundary static pressure. See Fig. 13.
The distributions of pressure ratio and isentropic efficiency
ofboth channel-wedge diffusers are shown in Figs. 14 and 15,
wherethose of other different types of diffusers are reproduced to
becompared together. The other different types are directly from
theauthors’ studies of Part I to III including a purely vaneless
dif-fuser. The “LSD Single Row” design has a solidity of 0.71 with
astagger of 19.58 deg which showed the best acceptable perform-ance
in the parametric analysis from Part I and II. The “LSDTandem”
design is the Tandem (A) with RCP¼ 0.3 in Part IIIwhich showed the
best performance in the tandem LSD paramet-ric analysis. Even
though the vane throat is choked at far earlier
than the design flow rate in Channel-wedge Optimal, the
designshowed the highest pressure ratio of 4.31 and the highest
isentropicefficiency of 81% due to the optimal combination of
design param-eters. However, for the application of the current
centrifugal com-pressor design duty, the optimal has to be
abandoned to move toChannel-wedge design to secure a reasonable
choke margin atdesign flow rate. At design flow rate, it showed the
highest effi-ciency of 81%, which was the same as that of
Channel-wedge Opti-mal, and the pressure ratio of 4.15. In terms of
the operating range,Channel-wedge Optimal may look inferior to
Channel-wedge, butagain the smallest flow on the map does not mean
a true stall/surgeflow, which would be one of restrictions that any
CFD study has.
An interesting result is that nearly the same level of the
pressurerise distribution was found in LSD Tandem in spite of a
slightdrop of maximum efficiency to 80% which value is still highly
ac-ceptable. Furthermore, it was found to provide wider range
ofoperation than Channel-wedge, as shown Fig. 16 (where the rangeof
operation is defined as the ratio of mass flow rate change
Table 2 Design Parameters of channel-wedge diffuser vanes
Version L/w3th AS AR34 2h 2u a3b NV
Channel-wedge Optimal 14.8 1.10 2.86 7.45 3.14 16.2
34Channel-wedge 11.0 0.84 2.68 8.81 3.60 18.0 29
Note: See Fig. 13 for design parameter definitions.
Fig. 13 Static pressure contours at design flow in channelwedge
diffuser vanes
Fig. 14 Pressure ratio characteristics of different
diffusertypes
Fig. 15 Efficiency characteristics of different diffuser
types
Fig. 16 Performance comparison at design flow for
differentdiffuser types
Journal of Turbomachinery NOVEMBER 2012, Vol. 134 / 061025-7
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
-
between maximum and minimum flow rates to maximum flowrate,
depending on the solution convergence when exit static pres-sure is
specified as a boundary condition.) The next lower per-formance was
found in the LSD Single Row which comes betweenthe vaneless
diffuser and LSD Tandem or Channel-wedge. It isworthy to note that
the maximum isentropic total-to-static effi-ciency of a single-row
LSD in a high-pressure (around 4.0 oftotal-to-static stage pressure
ratio) centrifugal compressor like thepresent machine is around 78%
at most. Of course, when the vo-lute is added, an additional
recovery of static pressure in the vo-lute will raise the
total-to-static efficiency.
In conclusion, the tandem LSD diffuser is recognized as one
ofattractive options in the selection of vaned diffusers for
bothdesign issues of the efficiency and the operating range.
References[1] Oh, J. S., and Agrawal, G. L., 2007, “Numerical
Investigation of Low Solidity
Vaned Diffuser Performance in a High-Pressure Centrifugal
Compressor—PartI: Influence of Vane Solidity,” ASME Paper No.
GT2007-27260.
[2] Oh, J. S., Buckley, Ch.W., and Agrawal, G. L., 2008,
“Numerical Investigationof Low Solidity Vaned Diffuser Performance
in a High-Pressure CentrifugalCompressor—Part II: Influence of Vane
Stagger,” ASME Paper No. GT2008-50178.
[3] Pampreen, R. C., 1972, “The use of Cascade Technology in
Centrifugal Com-pressor Vaned Diffuser Design,” Trans. ASME J. Eng.
Power, 94, pp.187–192.
[4] Senoo, Y., Hayami, H., and Ueki, H., 1983, “Low-Solidity
Tandem-CascadeDiffusers for Wide Flow Range Centrifugal Blowers,”
ASME Paper No.83-GT-3.
[5] Osborne, C., and Sorokes, J. M., 1988, “The Application of
Low Solidity Dif-fusers in Centrifugal Compressors,” Flows in
Non-Rotating TurbomachineryComponents, ASME FED, 69.
[6] Japikse, D., 1996, Centrifugal Compressor Design and
Performance, ConceptsETI, White River Junction, VT.
[7] Seleznev, K. P., and Galerkin, I. B., 1982, Centrifugal
Compressors, L:Mashi-nostroenie, Leningrad Division, Moscow,
Russia.
[8] Bandukwalla, P., 1988, “Diffuser Having Split Tandem Low
Solidity Vanes,”US Patent No. 4824325.
[9] Oh, J. S., and Ro, S. H., 2001, “Analysis of 8 Centrifugal
Compressor ImpellersUsing Two Different CFD Methods—Part I: Code
Validation,” ASME PaperNo. 2001-GT-326.
[10] Oh, J. S., 1998, “Numerical Investigation of Internal Flow
Field for ModifiedEckardt Backswept Impeller,” ASME Paper No.
98-GT-296.
061025-8 / Vol. 134, NOVEMBER 2012 Transactions of the ASME
Downloaded 28 Jan 2013 to 132.244.95.6. Redistribution subject
to ASME license or copyright; see
http://www.asme.org/terms/Terms_Use.cfm
http://dx.doi.org/10.1115/1.3445671
F1lF2F3F4T1F6F8F5F7F10F11F9xAF12T2F13F14F15F16B1B2B3B4B5B6B7B8B9B10