PARTIALLY VANED DIFFUSER WITH VARIABLE CROSS-SECTION FOR CENTRIFUGAL FANS
Tore Fischer Institute of Turbomachinery and Fluid Dynamics,
Leibniz Universität Hannover D-30167 Hannover, Germany
Sebastian Burgmann Chair of Fluid Mechanics
Bergische Universität Wuppertal D-42119 Wuppertal, Germany
Manuel Rudersdorf The Fuel Cell Research Center
ZBT GmbH D-47057 Duisburg, Germany
Joerg R. Seume Institute of Turbomachinery and Fluid Dynamics,
Leibniz Universität Hannover D-30167 Hannover, Germany
ABSTRACT The present research focuses on the efficiency improvement
at part-load of a centrifugal fan for a 30 kW fuel cell combined
heat and power (CHP) unit. For this purpose, the fan stage is
equipped with a partially vaned diffuser with a variable cross-
sectional area using a moving backplate.
The design and the performance of the partially vaned
diffuser with a variable cross-sectional area are described in
the first part of this paper. The performance results are com-
pared to measurements of the same centrifugal fan with a
vaneless diffuser carried out for the previous investigation. For
the second part, the influence of the variable cross-sectional
area on the diffuser flow field is investigated using optical PIV
(Particle Image Velocimetry) measurements and CFD (Compu-
tational Fluid Dynamics) simulations.
The combination of a variable cross-section, partially
vaned diffuser was able to achieve a 10 percent increase in
pressure ratio, a 5 percentage points increase in part-load effi-
ciency while maintaining the whole operating range of the
vaneless, constant cross section reference design.
INTRODUCTION This paper presents the second part of an ongoing investi-
gation on the performance improvement of a centrifugal fan for
the air supply of a 30 kW fuel cell system [1]. The objective of
this research is the improvement of the part-load operation of
the centrifugal fan by means of the variability of the cross-
sectional area of the radial diffuser and the volute. Through
improving the performance, the parasitic power consumption of
the fan can be reduced, resulting in an increased part-load
efficiency of the fuel cell system.
As a part of the transition process in the energy supply, the
demand for decentralized energy conversion is growing. Small
energy conversion units have to operate under variable power
demand at high efficiencies, but more importantly, they have to
be cost-effective in order to achieve customer acceptance.
Particularly small fuel cell systems show great potential for the
decentralized power and heat co-generation, with regards to the
decreasing heat demand of newly constructed and modernized
buildings. They achieve high electrical efficiencies and high
power-to-heat ratios for the whole operating range.
In contrast to automotive fuel cell applications, these
stationary systems are operated at moderate pressure levels
because of the stack size, and for this reason the power density
is less important. Consequently, radial fans are a suitable solu-
tion for the cathode air supply. They achieve high peak efficien-
cies at moderate pressure ratios, but more importantly, they are
a low-cost, state-of-the-art technology.
One major disadvantage of radial fans is the rapid
decreasing efficiency at off-design operation. This operating
behavior is a significant penalty for the overall efficiency of the
fuel cell system at part-load operation, as described among
others by Kulp et al. [2]. For this reason, the development of
simple and cost-effective performance stabilizing devices is of
increasing importance, not only with regard to stationary fuel
cells.
There are a number of measures and devices to improve the
part-load operation of turbomachines, as already presented in a
previous paper [1]. However, most of them are either very
Proceedings of ASME Turbo Expo 2017: Turbomachinery Technical Conference and Exposition GT2017
June 26-30, 2017, Charlotte, NC, USA
GT2017-63965
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complex and therefore usually used only for high-cost
applications, or they have drawbacks in terms of peak efficiency
reduction.
For these reasons, the presented approach focuses on a
simple variability of the diffuser and volute cross-sectional
areas by means of a movable backplate. This approach was first
presented by Lohmann [3] for the diffuser of a centrifugal fan,
albeit without considering the volute geometry necessary for a
real application. The approach has two advantages. First, the
aerodynamic design can be optimized for high peak efficiency,
e.g. due to the application of guide vanes. Secondly, the
actuation mechanism only has to move linearly and hence is of
reduced complexity.
The variability of the cross-sectional area stabilizes the
diffuser flow through the variability of the diffuser area ratio,
and hence variability of the static pressure rise. At low flow
rates, where the diffuser flow is highly tangential (flow angle
≥ 80 degree), the reduction of cross sectional area leads to an
increase of the radial velocity component. As a result, the flow
angle and the friction losses can be reduced.
AERODYNAMIC DESIGN The aerodynamic design of the centrifugal fan is performed
based on a combined analytical and numerical design pro-
cedure, as presented in detail by Fischer et al. [1]. An overview
of the design procedure is depicted in Fig. 1.
Boundary conditions
Impeller slip factorStodola [4]
Diffuser lossesEckert and Schnell [5]; Zahn [6]
Volute lossesEck [7]; Zahn [6]
Volute geometry
parameters
CFD simulation
Design procedure:
3D geometry
Ad
just
men
t of
loss
para
met
ers
Figure 1: Diffuser and volute design procedure [1]
The design approach combines 1D performance predictions
and 3D CFD simulations in an iterative loop. For the first step,
1D performance predictions based on published loss correla-
tions [4-7] are carried out. The initial geometry is then defined
based on the 1D performance prediction with an initial guess of
the loss parameters. In the following step, a 3D CFD simulation
of the initial geometry is used to adjust the loss parameters.
Based on these adjusted loss parameters, a new volute and
diffuser design is generated. This loop is repeated until the
analytical and the numerical solutions achieve sufficient
convergence.
For the present investigation the experimental and
numerical results of two aerodynamic designs are evaluated.
The first design, the reference design, consists of a vaneless
diffuser of 110 mm outlet diameter and a volute both with a
movable backplate. The second design consists of a partially
vaned diffuser with 4 mm vane height, and 100 mm outlet
diameter and a volute with reduced throat area, both with a
movable backplate. The important geometrical parameters of
both versions are summarized in Tab. 1.
Table 1: Geometrical parameters of the vaneless and the vaned diffuser Parameter Vaneless Vaned
Throat centroid radius RTT 84.7 mm 70.7 mm
Range of ATT/RTT 1.17 – 1.48 in 0.65 – 0.82 in
Diffuser outlet radius R3 55 mm 50 mm
Range of diffuser width b3 2 – 14 mm 4 – 12 mm
The centrifugal fan and the electric motor, used for the
proof of concept of the variable cross-sectional area, are
provided by Vorwerk Elektrowerke GmbH. A brief overview of
the fan performance data is given in Tab. 2. This fan is chosen
as cathode air supply for a 30 kW fuel cell combined heat and
power unit. The fuel cell stack consists of 100 cells. Each cell
has an active area of 672 cm². The fuel cell is operated at a
stoichiometry of two and the pressure loss of the whole stack is
approximately 7,500 Pa. Consequently, the rated mass flow rate
of the air supply is 0.05 kg/sec and the rated shaft power is
734 W.
Table 2: Performance data of the centrifugal fan
Impeller inlet radius R1 16.8 mm
Impeller outlet radius R2 28.5 mm
Blade width at impeller outlet b2 8 mm
Max. rotational speed 65,000 min-1
Max. electric input power 1,000 W
The intention of the partially vaned diffuser is the increase
of stage pressure ratio at part-load operation for the closed
position of 4 mm diffuser width. As a result of the variable
diffuser width, the flow range can be kept equal to that of the
vaneless diffuser. The vaned diffuser and volute combination
with variable cross-sectional area is illustrated in Fig. 2.
The housing backplate consists of an outer and an inner
backplate to allow the variability of the diffuser and volute
cross-sectional area. The outer backplate seals the fan housing
against the environment. The inner backplate is movable in
axial direction. The movement in the inward direction (closing)
is limited by the vane height (4 mm) of the partially vaned
diffuser. The movement in the outward direction (opening) is
limited by the outer backplate. This leads to a maximum axial
movement of 8 mm respectively variable diffuser widths be-
tween 4 mm (close position) and 12 mm (open position).
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Impeller
Guide vane
Volute
ω
b2
R2
b3
RTT
R1
R3Diffuser
Moving
backplate
Figure 2: Cross-sectional view of the vaned diffuser (closed position 4 mm width; open position 10 mm width)
The demonstrator fan stage is depicted in Fig. 3. The vaned
diffuser insert is milled from Plexiglas to enable the illumi-
nation of the guide vane passages for the PIV measurements.
Figure 3: Demonstrator fan stage with variable cross-section
METHODOLOGY The present project combines experimental and numerical
research. The experimental part is performed at the Fuel Cell
Research Centre (ZBT). The numerical part is performed at the
Institute of Turbomachinery and Fluid Dynamics (TFD) of the
University of Hannover.
Experimental Setup For the experimental investigation, steady-state thermo-
dynamic measurements are performed at the blower test rig of
the ZBT. In addition to these conventional measurements, laser
optical PIV measurements are performed, in order to gain a
deeper insight into the flow mechanisms and to validate the
numerical simulations.
The test chamber of the ZBT is designed and manufactured
in accordance to DIN EN ISO 5801: 2011-11 (Fans – Perfor-
mance testing using standardized airways). A throttle is located
at the outlet of the outlet pipe to control the back pressure. An
optical sensor is used to measure the rotational speed of the
motor. The electric power consumption of the motor is
measured by a wattmeter. The output signal of all measurement
devices is 0-10 V. The sampling frequency of 1 Hz is converted
by a 16-bit multichannel analog digital converter.
The static pressure is measured at the rear part of the
settling chamber using DS2 pressure transmitters with a
measurement range of 25,000 Pa (±125 Pa). For temperature
measurement, Pt100 thermocouples (±0.8 °C) are located in the
same measurement plane.
The volume flow rate is measured by means of a nozzle at
the inlet of the outlet pipe in accordance to DIN EN ISO 5167-
3:2003 (flow rate measurement of fluids by means of throttle-
devices in pipes with circular cross-section). The volume flow
calculation is based on the pressure difference between the
settling chamber and the nozzle. A DS2 pressure transmitter,
with a measurement range of 2,500 Pa (±20 Pa), is used to
measure this pressure difference. The accuracy of the volume
flow rate measurement is ±4% at design conditions. However,
at very low flow rates near surge, the volume flow rate
measurement uncertainty increases to ±30%. The fan overall
efficiency
elP
Vp (1)
is calculated by the ratio of aerodynamic power to electric
power consumption.
The experimental setup of the demonstrator fan stage
consists of three modules, as illustrated in Fig. 4. The first
component consists of the electric motor, the fan impeller, the
housing backplate, and the movable backplate of diffuser and
volute. The second component is the diffuser-volute housing,
which is equipped with individually laser cut quartz glass
windows for the PIV measurements. The third part is the inlet
pipe for fresh air and seeding intake.
Inlet pipe
Fan housing with
observation windows
Impeller and
housing backplate
Linear traverse
Electric
motor
Figure 4: Modular experimental setup of the demonstrator fan stage
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The experimental PIV setup is illustrated in Fig. 5. It
consists of a Litron Nano L dual cavity laser with 135 mJ pulse
energy at 15 Hz. The light sheet optics generates a light sheet of
less than 1 mm thickness from the laser beam in order to
illuminate the region of interest (ROI) inside the diffuser or
volute geometry. For each operating point, 400 images are
recorded with a frequency of 5 Hz, in order to achieve statistic
independence and hence convergence of mean values and
standard deviation. Each image is post-processed to improve
the quality, and then split into small interrogation windows of
64 x 64 pixels with 50 percent overlap including approximately
ten particle images each. Based on a cross-correlation analysis,
the flow velocity components can be evaluated for two
dimensions within the measurement plane.
Figure 5: Experimental setup for the PIV measurement
Numerical Setup The steady-state numerical simulations are performed with
the commercial software ANSYS CFX. The numerical model is
illustrated in Fig. 6. It includes a 360 degree model of the fan
stage and the complete test stand geometry in order to minimize
discretization errors due to geometrical simplifications.
Interface
Test rig
Plane of evaluation
at pressure side
Outlet
InletImpeller
VoluteDiffuser
Figure 6: Numerical model used for CFD simulations
All numerical domains (except for the volute) are meshed
using structured hexahedral grids. The volute domain is meshed
using an unstructured tetrahedral grid with additional prism
layers to resolve the near-wall velocity profiles. A summary of
the numerical grid quality criteria is given in Tab. 3.
Table 3: Numerical grid quality Minimum
orthogonal
Angle
Max. mesh
expansion
factor
Maximum
aspect ratio
Inlet Pipe 47.5° 3 537
Impeller 20° 18 407
Shroud cavity 22.2° 6 214
Vaned diffuser 30.9…35.7° 4…17 499…638
Volute 26…36° 20 199…213
Outlet pipe 41.6° 4 197
Test rig 55.8° 3 229
The boundary conditions are the total pressure and total
temperature at inlet (marked blue) and the mass flow rate at
outlet (marked red). For the impeller domain, a rotational speed
is set. The transition between domains with different frames of
reference is modeled with the frozen rotor approach. The
turbulence is modeled with the shear stress transport model
developed by Menter [8 and 9]. This turbulence model includes
a blending function between the use of Low-Reynolds and wall
function approaches. Thereby, the near-wall grid resolution can
be reduced significantly within the unimportant regions, e.g. the
test rig.
Figure 7: Numerical grid used for CFD simulations
The numerical grid is depicted in Fig. 7; it consists of ap-
proximately 25 million nodes. A grid convergence study [10]
showed that a further increase of the number of mesh nodes
does not lead to significant changes with regard to the target
values (efficiency and stage pressure ratio). The dimensionless
wall distances are below two for the diffuser and volute, which
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are the regions of interest (Tab. 4). The computational time of
the numerical simulations is approximately 1.5 million CPU
seconds per fan operating point. This leads to a run time of
approximately 42 hours on 72 cores with 2.9 GHz each.
Table 4: Numerical grid data Total number of grid nodes 24,246,672...26,189,819
Number of nodes (inlet pipe) 418,888
Number of nodes (impeller) 16,187,787
Number of nodes (shroud cavity) 901,600
Number of nodes (vaned diffuser) 4,120,336...5,884,527
Number of nodes (volute) 1,163,189...1,342,145
Number of nodes (outlet pipe) 170,560
Number of nodes (test rig) 243,920
Average y+ - value (inlet pipe) 0.46
Average y+ - value (impeller) 2.75
Average y+ - value (shroud cavity) 1.07
Average y+ - value (vaned diffuser) 0.31...0.36
Average y+ - value (volute) 0.61...1.06
Average y+ - value (outlet pipe) 1.02
Average y+ - value (test rig) 80.54
Validation of the Numerical Model For the validation of the numerical model both global and
local results are compared to the experimental measurements.
The global performance results are in accordance with
experimental results, as shown in Fig. 8. The comparison of
numerical simulations and experimental measurements for the
vaneless diffuser, with regard to the line of peak efficiency, is
shown in Fig. 8 (left). The deviation increases with rotational
speed, when the line of peak efficiency (white), which is
derived from numerical simulations, is compared to the area of
peak efficiency (red) of the experimental performance measure-
ments [1].
Figure 8: Comparison of experimental and numerical results of the vaneless diffuser (left) and the partially vaned diffuser (right)
In order to validate the numerical model for the part-load
performance prediction of the partially vaned diffuser, several
operating points between the surge line and the maximum flow
rate are compared to the experimental results (Fig 8, right). The
deviation between global numerical and experimental results
increases significantly for operating points at the maximum
flow rate at very low outlet pressures. The numerical error for
all operating points between peak efficiency and surge is
smaller than 4 percent regarding relative outlet pressure.
For the local validation of numerical results, it is assumed
that the prediction of the reverse flow inside the diffuser is of
greatest interest to the present research. For this reason, the
circumferentially averaged radial velocities are evaluated at two
radial positions (R = 35 and 45 mm). The velocity profile is
compared to the PIV results at three measurement planes at the
area right below the volute tongue. The results are depicted in
Fig. 9. It can be seen that the trend, as well as the occurrence of
backflow is sufficiently captured, although the deviation at the
outer radius (R = 45 mm) increases at low flow rates (OP A).
Figure 9: Circumferentially averaged radial velocity distribution from hub to shroud for three operating points (CFD versus PIV results)
RESULTS AND DISCUSSION The results of the experimental and numerical investigation
are split into two parts. Firstly, the global performance is
evaluated to demonstrate the effect of the cross-sectional
variability on fan performance. Secondly, local PIV and numeri-
cal results will be discussed, in order to explain some of the
physical phenomena which lead to the changes of fan perfor-
mance.
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Fan Performance The electric motor of the centrifugal fan is power con-
trolled. For this reason, the rotational speed depends upon the
electric power, the aerodynamically induced shaft torque, and
the mechanical losses.
For this reason, the performance results will be evaluated
with the common approach as outlet pressure versus stage mass
flow rate, and additionally as mass flow rate and outlet pressure
versus electric input power, in order to improve comparability.
The measured performance maps of the vaneless diffuser
are depicted in Fig. 10 for diffuser widths of 10 mm (left) and
4 mm (right). The small diffuser width of 4 mm leads to an
increase of stage outlet pressure at low mass flow rates. The
larger diffuser width of 10 mm leads to an increase of maximum
mass flow rate. The area of stage efficiencies larger than
40 percent is expanded and shifted to low flow rates for the
small diffuser width of 4 mm.
Figure 10: Measured performance maps of the fan stage with vaneless diffuser for 10 mm diffuser width (left) and 4 mm diffuser width (right)
The performance map for mass flow rate versus input
power confirms that the larger diffuser width of 10 mm in-
creases the maximum mass flow rate for all electrical input
powers. The evaluation of outlet pressure versus electrical input
power confirms that the small diffuser width of 4 mm leads to
an increased stage pressure ratio for all electrical input powers.
However, the effect of the variable cross-sectional area is
comparatively small for the vaneless diffuser. The approach for
the second design is to improve stage efficiency and pressure
ratio. For this purpose, the diffuser is equipped with guide
vanes. This common approach increases the static pressure
recovery of the diffuser, though it decreases the stable operating
range.
However, due to the variable cross-sectional area the stable
operating range can be extended and efficiency increased again.
As a result, no reduction of maximum flow rate and only minor
penalties of high-flow efficiencies have to be accepted.
Figure 11: Measured performance maps of the fan stage with vaned diffuser for 10 mm diffuser width (left) and 4 mm diffuser width (right)
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The performance maps of the partially vaned diffuser are
shown in Fig. 11 for the 10 mm diffuser width (left) and the
4 mm diffuser width (right). With the application of guide
vanes, the maximum stage pressure ratio is increased by over
10 percent, compared to the vaneless diffuser. Furthermore, the
peak efficiency is increased by 5 percentage points for the
closed position of 4 mm diffuser width, where the partially
vaned diffuser becomes a completely vaned diffuser.
For the small diffuser width of 4 mm, the stage outlet
pressure is increased by up to 15 percent at low mass flow rates.
The large diffuser width of 10 mm, on the other hand, achieves
the same maximum mass flow rate as the vaneless diffuser.
The performance map for mass flow rate versus input
power confirms that the larger diffuser width of 10 mm
increases the maximum mass flow rate for all electrical input
powers. The evaluation of outlet pressure versus electrical input
power confirms that the small diffuser width of 4 mm leads to a
major increase of stage pressure ratio for all electrical input
powers.
Local Effects For the evaluation of local flow phenomena, PIV measure-
ments are performed in order to explain the differences in
centrifugal fan performance with the vaneless and vaned
diffuser. For this purpose three measurement planes (Fig. 12)
inside the diffuser at an area right below the volute tongue
(ROI) are discussed below. The first plane (h1) is located near
the diffuser hub. The second plane (h5) is located at 50 percent
diffuser width. The third plane (h7) is located near the diffuser
shroud.
Figure 12: Region of interest (ROI), planes of evaluation, and operating points of the PIV measurements
The results of both diffusers are evaluated for three differ-
ent operating points. The first operating point A is near surge.
The second operating point B is at peak efficiency. The third
operating point C is close to the maximum flow rate. All three
operating points are illustrated in Fig. 12.
The diffuser flow fields near surge are illustrated in Fig. 13
by means of velocity contour plots, for both the vaneless and
the vaned diffuser. PIV measured flow fields for 10 mm diffuser
width are shown on the left and the flow fields for 4 mm
diffuser width are shown on the right.
The inlet Mach number of all diffusers is between 0.2 and
0.3, and the diffuser inlet radius to width ratio is between 0.1
and 0.3. According to Senoo and Kinoshita [11] this leads to
critical inlet flow angles between 80 (10 mm width) and
85 degrees (4 mm width) for the vaneless diffuser.
The comparison top-down of the 10 mm diffuser width (left
column) shows that reverse flow occurs near the hub and the
centre for both the vaneless and the vaned diffuser. At the
shroud, reverse flow only occurs for the vaned diffuser.
Figure 13: PIV results of the vaneless and vaned diffuser, three spanwise positions, operating point near surge (10 mm width left and 4 mm width right)
The comparison of the flow fields of 10 mm (left) and
4 mm (right) diffuser width confirms that the reduced width
stabilizes the diffuser flow field. This agrees with the described
increase of critical flow angle due to the reduced ratio of
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diffuser inlet radius to width. As a result, no reverse flows occur
for both the vaneless and the vaned diffuser.
However, the vaned diffuser achieves stronger flow decel-
eration due to the flow guidance. For this reason, the vaned
diffuser has a higher stage pressure ratio near surge than the
vaneless diffuser.
The PIV results at peak efficiency are depicted in Fig. 14.
The flow fields of the vaneless diffuser at open position (10 mm
width) show an area of reverse flow only near the hub. This
result agrees with the results from the numerical simulation,
shown previously.
The flow fields of the vaned diffuser at open position
reveal that no throughflow takes place at the vaned part of the
diffuser. The flow field of the vaneless part shows a stable flow
field. Consequently, the vanes cause additional losses at design
flow condition.
Figure 14: PIV results of the vaneless and vaned diffuser, three spanwise positions, operating point at peak efficiency (10 mm width left and 4 mm width right)
The PIV results of the diffuser at closed (4 mm) position
(Fig. 14, right) show a stable flow field for both the vaneless
and the vaned diffuser. This agrees with the increase of stage
peak efficiency shown previously by means of the stage perfor-
mance maps (Fig. 10 and 11).
The diffuser flow fields for an operating point close to the
maximum flow rate are illustrated in Fig. 15 by means of PIV
measurements. The results of the vaneless diffuser at open posi-
tion (10 mm, left) illustrate the stable diffuser operation. The
results of the vaned diffuser at open position (left) show low
flow velocities in the vaned part of the diffuser whereas the
flow field in the vaneless part is stable. As a result, the vaneless
diffuser achieves a higher efficiency as the partially vaned
diffuser, at high flow rates. This agrees with the performance
measurements shown in Fig. 10 and 11.
The results of both diffusers at closed position (4 mm,
right) highlight, that only a minor flow deceleration can be
achieved at this operating point. In addition, the PIV results of
the vaned diffuser show a large misalignment of the flow and
the vanes. This leads to flow separation at the suction side of
the guide vanes and as a result to increased aerodynamic losses.
Figure 15: PIV results of the vaneless and vaned diffuser, three spanwise positions, operating point close to the maximum flow rate (10 mm width left and 4 mm width right)
This also agrees with the performance measurements. The
efficiency of the large diffuser width (10 mm) increases with
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mass flow rate and the efficiency of the small diffuser width
(4 mm) decreases with mass flow rate.
For the next part of the local performance evaluation, the
volute and the vaned diffuser will be compared separately based
on CFD simulations. In addition to operating point A (480 W
electrical input power), which was already discussed based
upon PIV results, a second operating point at an electrical input
power of 220 W will be evaluated. Both operating points are
illustrated in Fig. 16 for the open and the closed position of the
vaned diffuser.
Figure 16: Operating points near surge for the evaluation of local CFD results
For the comparison two dimensionless performance para-
meters are used. The first parameter is the static pressure rise
coefficient
inin
inout
pp
ppCP
,0
(2)
This parameter is calculated from the difference of outlet static
pressure pout and inlet static pressure pin normalized by the
dynamic pressure at the domain inlet. This parameter describes
the amount of kinetic energy which is converted to potential
energy within the control volume.
The second parameter is the total pressure loss coefficient
inin
outin
pp
ppCPL
,0
,0,0 (3)
This parameter is calculated from the difference of inlet total
pressure p0,in and outlet total pressure p0,out normalized by the
dynamic pressure at the domain inlet. This parameter describes
the amount of kinetic energy which is lost within the control
volume.
Both dimensionless performance parameters will be evalu-
ated at the diffuser and the volute outlet normalized by the
kinetic energy at diffuser inlet. The results for the two operating
points for both diffuser widths are shown in Fig. 17. The static
pressure rise coefficient confirms the previous results. The
vaned diffuser at closed position (4 mm) achieves a static
pressure rise twice as high as the vaned diffuser at open
position. The static pressure rise of the volute is approximately
10 percent for both diffuser widths.
At open position, the total pressure loss coefficient
illustrates the additional losses of the vaned diffuser at low-
flow. Almost 60 percent of the kinetic energy at the diffuser
inlet is lost within the diffuser domain. Within the volute, an
additional 10 percent of the kinetic energy is lost. As a result,
almost 70 percent of the kinetic energy is lost and only
30 percent is converted to static pressure.
At closed position (4 mm), the total pressure loss within the
diffuser is only 30 percent. Within the volute, an additional
20 percent of kinetic energy is lost. As a result, almost
50 percent of the kinetic energy is lost and 50 percent are
converted to static pressure.
Figure 17: Static pressure rise coefficient (left) and total pressure loss coefficient (right) at diffuser (top) and volute outlet (bottom) at operating points near surge
These results confirm the assumptions discussed based on
the PIV results. In order to identify the locations of high-losses
within diffuser and volute, the pressure loss coefficient is
depicted in Fig 18 by means of contour plots of the open
(10 mm, top) and the closed position (4 mm, bottom).
The illustrations of the total pressure loss agree with the
previous results, as the magnitude is obviously higher at open
position (10 mm). The area of high losses is located in the
vaned part of the diffuser and the losses in the vaneless part are
equal to the losses of the vaned diffuser at closed position
(4 mm).
For the vaned diffuser at closed position (4 mm), there are
also areas of high total pressure loss in the vicinity of the vanes.
However, the magnitude is approximately 20 to 30 percent
lower and the area is significantly smaller, compared to the
open position (10 mm). This also agrees with the assumptions
made previously based on PIV results.
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Additionally, the CFD results first reveal that the total
pressure loss appears to be increased in some single vane
passages. These passages are marked by red arrows in Fig. 18
(bottom) and will be also discussed below.
Figure 18: Total pressure loss coefficient at open (top) and closed position (bottom) at an operating point near surge (480 W)
The static pressure rise of the vaned diffuser at open
(10 mm, top) and closed (4 mm, bottom) position is illustrated
in Fig. 19, by means of contour plots. The overall magnitude
confirms previous results: the vaned diffuser at closed position
(4 mm) achieves a higher static pressure rise near surge. At
open position (10 mm), the major static pressure rise takes
place in the vaneless part of the diffuser.
In addition, it can be seen that the pressure rise coefficient
in the volute is uniform for both diffuser positions. This shows
that the volute of the vaned diffuser is sufficiently dimensioned
for the part-load operation.
The previously mentioned vane passages with increased
losses are marked by white arrows in Fig 19 (bottom). While
these passages provide no pressure rise the static pressure rise
coefficient remains constant. Based on this result, it can be
assumed that reverse flow occurs in these passages. As a result
pressurized air from the diffuser outlet flows upstream and
increases the static pressure in the vane passage.
The occurrence of reverse flow in individual vane passages
is an indication of diffuser stall onset. Stall is an unsteady flow
phenomenon. This may explain the increasing local deviation
between PIV and CFD results shown above (Fig. 9) since
unsteady flows cannot be captured accurately by steady-state
numerical simulations.
Figure 19: Static pressure rise coefficient at open (top) and closed position (bottom) at an operating point near surge (480 W)
However, the occurrence of reverse flow can be identified
by evaluation of the radial velocity component, despite local
inaccuracy in the stall region. For this purpose, the radial
velocity component is depicted in Fig. 20 for the vaned (top)
and vanless (center) part of the diffuser at open position and for
the vaned diffuser at closed position (bottom). Additionally,
velocity vectors of the absolute velocity are added to the
contour plots in order to visualize the flow direction.
The contour plot of the closed position (bottom) shows that
one of the vane passages is completely blocked by reverse flow
(black arrow) and the other passage is partially blocked by
reverse flow (yellow arrow). This confirms the assumption of
stall onset. The counterclockwise adjacent vane shows flow
separation at the suction side and the clockwise adjacent vane
shows no flow separation. This behavior is caused by the flow
displacement due to the blocked passage which changes the
angle of attack of the adjacent vanes. This well-known
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phenomenon has been described by various authors, e.g. a good
description of cascade stall phenomena can be found in [12].
The contour plot of the open position (10 mm) in Fig. 20
(top and center) illustrates that the vaned part of the diffuser is
blocked by reverse flow between 90 and 360 degrees. Only in
the section between the volute tongue (0 degrees) and 90 de-
grees, the radial velocity component is positive and hence no
reverse flow occurs (Fig. 20, top). For the vaneless part at open
position (Fig. 20, center) the diffuser flow is stable with regard
to the radial velocity component.
Figure 20: Radial velocity at open (top and center) and closed position (bottom) at an operating point near surge (480 W)
For the last part of the local performance evaluation, the
performance of the vaned diffuser at open position (10 mm) will
be compared for operating points near surge and close to the
maximum mass flow rate. All operating points are illustrated in
Fig. 21 for two electrical input powers of 220 and 480 W.
For the comparison, the dimensionless performance para-
meters CP (Eq. 2) and CPL (Eq. 3) are used. The results of the
two operating points for both electrical input powers are shown
in Fig. 22.
Figure 21: Operating points at low-flow (LF) and high-flow (HF) for the evaluation of local CFD results
Figure 22: Static pressure rise coefficient (left) and total pressure loss coefficient (right) at diffuser (top) and volute outlet (bottom) at low-flow (LF) and high-flow (HF) operating points (10 mm diffuser width)
The vaned diffuser at open position achieves a static pres-
sure rise coefficient of 50 percent at high-flow, compared to
20 percent at low-flow. At the volute outlet the static pressure
rise coefficient drops to approximately 30 percent at high-flow.
This drop of pressure rise coefficient indicates that flow
acceleration takes place in the volute. As a result, the static
0°
90°
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pressure rise is partially lost and hence total pressure losses in-
crease.
This increase of total pressure loss is illustrated in Fig. 22
(right). At high-flow, the total pressure loss coefficient of the
diffuser is below 20 percent. Between diffuser and volute outlet
the value of the total pressure loss coefficient increases to
40 percent. The increase of loss agrees with the decrease of
static pressure rise at the volute outlet.
In order to illustrate the locations of static pressure rise and
total pressure loss, both coefficients are depicted in Fig. 23 by
means of contour plots for an operating point close to the
maximum mass flow rate. The total pressure loss coefficient
implies that a large part of the losses occur downstream of the
volute throat area. The size of this area defines the volute
design flow rate. A small throat area is suitable for low-flow
applications and a large throat area is suitable for high-flow
applications.
In this case the volute is designed for low-flow operation.
Consequently, the throat area is designed too small for
operating points close to the maximum mass flow rate. As a
result, the flow is accelerated and hence the static pressure rise
coefficient decreases downstream of the throat area (Fig. 23,
bottom).
Figure 23: Total pressure loss coefficient (top) and static pressure rise coefficient (bottom) at open position for an operating point close to the maximum mass flow rate (480 W)
CONCLUSIONS The range extension of a centrifugal fan by means of a
partially vaned diffuser with variable cross-sectional area is
achieved and physically explained based upon numerical and
experimental data. The results obtained by experimental and
optical PIV measurements were compared to those of numerical
simulations. The trend of the diffuser flow distribution, as well
as the occurrence of backflow was sufficiently captured by
numerical simulations. It was found that the minor increase of
local deviation between PIV and CFD results towards low mass
flow was caused by stall onset at operating points near surge.
The performance measurements prove that the approach of
variable cross-sectional area is suitable for range extension.
Both the optical measurements and the CFD simulations
confirm the potential of the cross-sectional variability con-
cerning fan performance improvement at part-load operation.
The original design with a vaneless diffuser achieved a
higher efficiency at high mass flow rates compared to the new
partially vaned diffuser. However, the experimental perfor-
mance measurements confirmed that the stable operating range
of the partially vaned diffuser was extended due to the variable
cross-sectional area. As a result, the partially vaned diffuser
achieved the same flow range as the vaneless diffuser.
Due to the application of guide vanes, the maximum stage
pressure ratio was increased by over 10 percent, compared to
the original design with a vaneless diffuser. Furthermore, the
peak efficiency and the part-load efficiency were increased by
over 5 percentage points for the closed position of 4 mm dif-
fuser width.
In summary, the combination of a variable cross-section,
partially vaned diffuser was able to achieve a 10 percent in-
crease in pressure ratio, a 5 percentage points increase in part-
load efficiency while maintaining the whole operating range of
the vaneless, constant cross section reference design.
ACKNOWLEDGMENTS This paper presents results of the IGF research project
18100N of the IUTA and the FVV, which is funded by the AiF
within the program for the founding of industrial collaborate
research (IGF) of the BMWi based on a resolution of the
German Bundestag.
The authors would like to acknowledge the contribution of
Mr. Martin Meggle of Vorwerk Elektrowerke GmbH for
providing the centrifugal impeller and the electric motor. The
authors also gratefully acknowledge the guidance of the
industrial partners of the IGF Kleingebläse für Brennstoffzellen
committee. Further thanks are due to the Regionales Rechen-
zentrum für Niedersachsen (RRZN) and Institut für Werkzeug-
lose Fertigungsverfahren (IWF), University Duisburg-Essen.
Furthermore we thank Mr. Alexander Roos and Ms. Lara Baune
for their dedicated contributions to this project. Finally, we
acknowledge the valuable suggestions of the anonymous re-
viewers.
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NOMENCLATURE A area
b width
CP pressure rise coefficient
CPL total pressure loss coefficient
p static pressure
p0 total pressure
P power
R radius
volume flow rate
Δp pressure difference
η overall efficiency
SUBSCRIPTS 1 fan inlet
2 fan outlet / diffuser inlet
3 diffuser outlet
el electrical
in inlet
out outlet
TT volute throat section
ABBREVIATIONS CFD Computational Fluid Dynamics
CHP combined heat and power
CPU Central Processing Unit
HF high-flow
LF low-flow
OP operating point
PIV Particle Image Velocimetry
ROI region of interest
TFD Institute of Turbomachinery and Fluid Dynamics
ZBT the Fuel Cell Research Centre
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