Partially Vaned Diffuser With Variable Cross-Section for ... · PARTIALLY VANED DIFFUSER WITH VARIABLE CROSS-SECTION FOR CENTRIFUGAL FANS . Tore Fischer Institute of Turbomachinery

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

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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).



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



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



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.



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)



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



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



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.



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



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°



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.



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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the ASME Turbo Expo 2016, 13-17 June 2016,

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Partially Vaned Diffuser With Variable Cross-Section for ... · PARTIALLY VANED DIFFUSER WITH VARIABLE CROSS-SECTION FOR CENTRIFUGAL FANS . Tore Fischer Institute of Turbomachinery