PROCEEDINGS of the
23rd International Congress on Acoustics
9 to 13 September 2019 in Aachen, Germany

Aeroacoustic simulation and experimental validation of soundemission of an axial fan applied in a heat pump
Andreas LUCIUS1; Marc SCHNEIDER1; Stefan SCHWEITZER-DE BORTOLI2, Tom GERHARD3,
Thomas GEYER41

ebm-papst Mulfingen, Germany
2
3

4

Vaillant, Germany

former Vaillant, Germay

Brandenburg Universtity of Technology, Germany

ABSTRACTThe reduction of noise emission of heat pumps is a major development target for the appliance manufacturer.
An important source of sound is the fan. The fan is operating at disturbed inflow conditions due to installationin the appliance. Additionally small scale turbulence generated by the heat exchanger further increases fannoise emission. Noise emission may be reduced via improved installation or specific fan design. The aim of
the project was to measure flow and turbulence characteristics inside the heat pump via constant temperature
anemometry as well as noise emission. The acquired data set was then used for validation of CFD and CAA
simulations. The effect of inflow turbulence was analyzed for configurations with and without heat exchanger.
Hot wire measurements show increased average turbulence levels of 20 percent downstream the heatexchanger. In a second step URANS simulations were conducted for both configurations. The heat exchanger
was modelled as porous medium. Turbulence properties could be matched with experiments with reasonableboundary conditions at the heat exchangers outlet. There is some deviation in flow rate with URANS.
Increased tonal noise for the configuration with heat exchanger is predicted but the levels are in error of up to
10 dB.
Keywords: fan noise, heat pump, turbulence, CAA

1. INTRODUCTION
The reduction of CO 2 emissions is one of the most important global challenges. In Germany
conventional heating with gas or oil dominates the installed heating systems. An alternative to theusage of fossil fuels for heating are heat pumps. These machines use electrical power to tra nsfer
environmental heat to useable heat. If the electrical power is generated from renewable sources theoperation is CO 2 neutral. The increasing number of installed heat pumps in human environment likegardens raises the importance of noise emissions. The axial fan is one of the key components of a heatpump. In addition to the compressor the fan is the most important source of noise. This paper analyzes
the influence of the heat exchanger as turbulence generator on fan noise emission. The heat exchanger
is expected to generate small scale turbulence. On the other hand it acts as a fl ow straightener. Large
scale vortical structures are broken and flow homogeneity is increased. The present paper analyzes theinfluence of the heat exchanger on fan noise emission as well as its influence on turbulent inflow
conditions.
Disturbed turbulent inflow conditions are well known to increase fan noise (1). Turbulent
ingestion noise on airfoils has been the topic of numerous research papers in the past. Recent studie s
on increased fan noise due to inflow disturbance are e.g. the work of Zenger et. al. (2). The influenceof the heat exchanger in an automotive cooling fan was analyzed by Rynell et. al. (3). He found, thatremoving the heat exchanger in his setup increases fan noise. This is in line with the work of Lucius etal. (4), who showed dramatic increase for a rectangular box as inflow disturbance. The heat exchanger
1

Andreas.Lucius@de.ebmpapst.com

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itself generates small scale turbulence due to flow separation at the cooling tubes. But additionally thenarrow lamellas act as a flow straightener. Large scale turbulent structures as corner vortices from thebox are broken by the heat exchanger. The aim of the current study is to analyze the effect of inflow
turbulence on fan noise in a heat pump. There is only limited knowledge of turbulence properties of theflow after a heat exchanger. Length scales and turbulence intensity are the key parameters that
determine the leading edge noise. The work of Czwielong et. al. (5) shows hot wire measurements after
different heat exchanger configurations. These measurements were done without fan at the leading
position. The measurements show high turbulence intensity and anisotropy. The flow is accelerated
from rectangular heat exchanger cross section to the fan housing, which leads to vortex stretching andthe high level of anisotropy reported.
In the present paper hot wire measurements were done inside a heat pump. Two configurations wereanalyzed: The base configuration and a second setup without heat exchanger. Configurations are
displayed in Figure 1.

Y
Z

Figure 1. Configurations analyzed, left base configuration, right without heat exchanger (HEX)

2. Experimental results
2.1

Experimental setup and acoustic results

In order to analyze the effect of inflow turbulence due to the heat exchanger the two configurations
were measured in different test rigs. First of all acoustic and aerodynamic performance was measured
in a test rig according to DIN EN ISO 5801. The test rig is an anechoic room with reflecting bottom.
Pressure side and suction side are separated with a wall. The heat pump is mounted at the chamber wall.
Tape was used to prevent any airflow leakage between inlet and outlet of the heat pump. The heat pump
was lifted from the bottom such that fan axis is 1.5 m above the reflecting bottom. Additional throttling
was applied to see the influence of the operational point on the acoustic performance. Sound power
was measured on suction and pressure side with the enveloping surface method. Results shown in thispaper are focused on inlet sound power. Figure 2 shows the suction side overall sound power for thetwo configurations. The configuration with heat exchanger (HEX) shows up to 1 dB higher soundpower. This is not in line with the results of Rynell et al. (3). The main reason for this behavior is
attributed to the asymmetric disturbed inflow conditions even with the heat exchanger installed ( seeFigure 5). The inverter marked in blue is a very narrow disturbance, which has probably more effectthan flow separation from the casing if the heat exchanger is removed.
There is some dependence of overall sound power on flow rate with steep increase at low flow rates.
The fan installed in the heat pump with heat exchanger operates near the point of minimal noise
emission. The additional pressure loss of the heat exchanger results in a reduction of flow rate of 18 %at free blowing conditions.
Sound power 3 rd octave spectra are shown in the right hand side of Figure 2. Both configurations are
displayed at the reference flow rate of the base configuration. For the configuration without heatexchanger the sound spectrum at free blowing conditions is additionally shown. At the same flow rate
there is an increase of broad band noise of 1 to 2 dB for frequencies up to 2 kHz in case of heatexchanger installed. At higher frequencies acoustic damping of the heat exchanger reduces inlet soundpower. With increased flow rate noise emission at nearly all frequencies is 1 to 2 dB higher.

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Figure 2. Overall sound power performance curve and 3rd octave spectra
2.2

Detailed measurements of flow velocity

Flow velocity inside the heat pump was measured with Constant Temperature Anemometry (CTA).
The measurements were conducted outside the performance test rig. In this measurement setup it was
not possible to measure or control the air flow rate. Flow rate is different for both configurations asshown in section 2.1. A DANTEC 55P64 probe was applied to measure flow in fan axial direction andvertical direction according to Figure 1. The probe was mounted on a 2D automatic traverse system.
Velocity was measured for 12 s in each point, first 2 seconds are skipped in evaluation due to potential
probe vibration after traversing. The sampling frequency was 25.6 kHz, spectra are generated fromblocks with 16384 points. A rectangular grid of 6 x 6 points with a spacing of 40 mm covers little morethan one quadrant of the fan circumference. The installation is asymmetric mainly due to t he inverterbox (see Figure 5). All measurements consider the air flow path only, the cooling circuit was not in
operation. The fan installed is a 3 bladed axial fan of 450 mm diameter. The fan is directly mounted on
the EC motor running at 620 revolutions per minute.
Two measurement planes at different distances to the fan are measured to get an impression on
turbulent dissipation. The measurement planes are located at 40 and 80 mm downstream the heatexchanger. All diagrams in this paper concerning detailed flow analysis refer to the 80 mm position.
The axial distance between fan leading edge and heat exchanger is 158 mm. A comparison of measured
axial velocities is shown in Figure 3. As expected flow velocities are higher for the configuration
without heat exchanger. However the difference is higher than 21 % expected from flowrate. The flow
field is more homogeneous for the base configuration. This is especially true for the border points. In
addition there is a radial distribution with low velocity region near the impeller hub for configuration 2.
Turbulence intensities for both configurations are shown in the lower part of Figure 3. The turbulence
intensity displayed is normalized with the local mean axial velocity. The heat exchanger leads to a
virtual increase of turbulence levels from near zero to 17% in the core region. Near the borders
turbulence levels are higher for the configuration without heat exchanger. This is attributed to flow
separation from the casing. This can have dramatic influence on fan acoustics (4). Velocity
fluctuations are quite similar for both measured directions, so only the axial velocity is shown in thispaper.

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Figure 3. Measured axial velocity (top) and turbulence intensity (bottom) for both configurations
The influence of the heat exchanger on turbulent spectra is shown in Figure 4. The curve shows
velocity spectra of the axial velocity for all measured points 80 mm downstream the heat exchanger.
For the base configuration all spectra coincide within a narrow band. There are two tonal components
in the spectrum. One peak is at 33 Hz which is the blade passing frequency. The other peak at 5 6 Hz
corresponds to vortex shedding at the heat exchanger tubes. This peak corresponds to a Strouhal
number of 0.23 with mean axial flow velocity and tube diameter as length scale. Configuration 2
without heat exchanger shows quite different behavior. There are some points with very low
fluctuations, these points are located near the rotation axis. The other group of points contains large
amount of kinetic energy especially at low frequencies. These are probes located close to the boundaryedges. In addition to kinetic energy, turbulent length scales are evaluated from the autocorrelation ofvelocity fluctuations. Measured and simulated length scales are discussed in section 3.2.

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Figure 4. Measured turbulent velocity spectra for both configurations

3. CFD setup and results
3.1 URANS simulations
CFD simulations were conducted with the general purpose CFD code STAR-CCM+ 12.06. The k-H
realizable turbulence model was chosen. All simulations were done with compressible air, as theacoustics is taken into account. The boundary conditions are chosen to obtain the free blowing
operating point (zero pressure outlet and stagnation inlet). For the transient simulations non-reflecting
boundary conditions were applied. Suction side and pressure side are separated in the modelledgeometry as shown in Figure 5. The heat exchanger is modelled as a porous medium (displayed inyellow). All simulations start with a stationary Multi-Reference-Frame simulation with frozen rotor
position. These simulations lead to large underestimation of flow rate in range of 30 %. Agreementwith measurements is much better if transient fan rotation is included. Temporal resolution is 100 steps
per pitch with a second order discretization scheme. Total number of cells for the URANS computation
is 19 million polyhedrals. The mesh is identical for both configurations. The boundary layer isresolved with 12 prism layers on the fan blade and on the cylindrical part of the casing. Integral values
show negligible differences between a coarser mesh containing 13 million cells. The aeroacousticpressure is determined with the FW-H solver of STAR-CCM+. The far field sound pressure at 2
microphone locations on the suction side is calculated propagating from a permeable emission surface
displayed in green in Figure 5.

inverter

FW-H

fan

surface
HEX

Figure 5. CFD simulation models, left heat pump, right periodic HEX model

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3.2
Heat exchanger calibration
The heat exchanger was modelled as a porous medium. For calibration of the model coefficients aperiodic model of the detailed heat exchanger geometry (tubes and lamellas) was simulated. Figure 5
shows the simulated geometry. Cooling lamellas were simplified as flat plates. These simulations wereused to check pressure loss depending on flow rate and turbulence properties downstream the heatexchangers. The simulations were run unsteady due to vortex shedding at the cooling tubes. A stable
solution from steady state RANS could not be achieved.
The obtained pressure loss of 9 Pa is 30 % lower compared to the pressure rise of the fan alone at thesame operating point. This is attributed to the simplified geometry. For the momentum equations in the
porous region only the inertial term was used. Pressure loss coefficient was set according to fan
pressure rise. Turbulence intensity Tu and length scale / are evaluated from the CFD with theassumption of isotropic turbulence. In equation (1) Uz is the axial velocity and k the turbulent kinetic
energy. The length scale is evaluated from kinetic energy and dissipation rate H.
ටʹ ݇͵ܶ‫ ݑ‬ൌ 
ȁܷ௭ ȁ

(1)

݇ ଷȀଶ
߳

(2)

Ȧ ൌ ͲǤͷͶ

Figure 6 shows the averaged CFD results in comparison to the correlation of Roach (6) for parallelrods with the tube diameter D of 7 mm used as length. Correlation and simulation results for theperiodic model coincide quite well. For turbulence intensity there is some difference at small distance s
where the correlation is not valid. The graphs also show the average measured value of turbulent
length scale and turbulence intensity for both measurement planes. These values correspond quite well
to CFD and Roach correlation. The experiments indicate higher turbulent length scales than expected
at the shorter distance. This may be attributed to the short distance to the heat exchanger which is
below the data limit of 10 D in (6). As shown in section 2.2 measured kinetic energy also contains
contribution due to impeller rotation. This content is not broad and for that reason is not included in themodelled kinetic energy. For that reason measured kinetic energy is expected to be higher than from
CFD results. CFD data of the periodic model will now be used as input for the porous medium in thefull simulation. Fitting turbulence is quite challenging. The boundary conditions in terms of lengthscale and turbulence intensity are specified at heat exchanger outlet. Turbulence intensity dissipate swith distance depending on the length scale applied. Initial RANS simulations of decaying turbulencefrom turbulent inflow conditions were performed. It was not possible to fit both properties (/ and Tu)
at measured locations. RANS with turbulent inlet boundary lead to a decay similar to Roach, which hasstronger dissipation compared to the periodic resolved HEX model. A variation of inlet parameters wasdone in order to fit turbulence intensity. The best fit was achieved with considerable larger lengthscales. The chosen turbulent properties at HEX outlet are Tu = 30 % and / = 4.5 mm. These values areused as input for the full simulation of the base configuration with heat exchanger.

Figure 6. Turbulence properties from experiments and simulations

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3.3 URANS Results
In this section URANS simulation results are discussed. The results correspond to temporal meanvalues averaged over 5 out of 8 rotor revolutions. Table 1 shows integral performance in comparison tomeasured values. Absolute flow rate Q v is ca. 10% smaller compared to the experiments for bothconfiguration, but the increase of 21 % in flow rate between configuration 1 and 2 is well predicted.
Simulated velocity and turbulence fields are shown in Figure 7. Velocities are quite well predicted forboth configurations. However there is an offset due to error in flow rate. In addition flow field is morehomogeneous in simulations compared to measured results. With proper settings for the porous
medium good agreement of turbulence levels is achieved (see Figure 6). Simulations show flow
separation from the casing for configuration 2. Velocity defects and high levels of turbulence similar tomeasurements are found in the outer region of the measurement planes.
Table 1. URANS results vs experiments
configuration

nr. of cells

1

1.9 1072

7

1.9 10CFD

experiment

Qv / Qv Base Exp.

Qv / Qv Base Exp.

0.889

1.0

1.066 (+19.7%)

1.215 (+21.5%)

Figure 7. Time averaged URANS results, velocity and turbulence for both configurations
Finally aeroacoustic results are discussed. Figure 8 displays acoustic spectra for bothconfigurations at free blowing conditions. The microphone positions are at 1 m distance to the1857appliance on fan axis and 3.4 m off axis respectively. Simulation and experiments are displayed withthe same frequency resolution of 5.5 Hz. CAA only provides tonal components at BPF and its
harmonics. For tone 1 (33 Hz) the frequency is too low for the acoustic test rig. The correct trend
between both configurations is predicted for tones 2 and 3. Configuration 1 shows higher tonal noise,with stronger difference at tone 3. The trend is not correct for tone 4. Tones higher than 5 th order are
much too low which is attributed to lack of temporal resolution in the simul ation. Absolute levels show
deviations in the range of 10 dB between measurements and simulations. Even though the simulation
time is rather short, CAA results are not expected to improve with longer physical time. Time signals
show smooth periodic behavior with very little variation in time.

Figure 8. Aeroacoustic simulation results vs. experiments

4. Conclusion
In this paper the effect of heat exchanger in a heat pump on fan noise was investigated.
Measurements and simulations show high levels of turbulence in range of 20 % generated by the heatexchanger. Additional turbulence leads to increased fan noise emission of 1 to 2 dB in 3 rd octave
spectrum. Measured velocity and turbulence are in good agreement with simulations. For the heatexchanger it is important to choose proper boundary conditions to achieve correct turbulence levels.
Tuning can be done with auxiliary simulations of detailed heat exchanger geometry. The influence ofthe heat exchanger on tonal noise was simulated via FW-H based on URANS. The trend between
configurations is correctly predicted for most of the tones up to 5 th order. The error of absolute levels
is up to 10 dB for single tones. For an evaluation of the broad band noise emission tur bulence
generators are needed as the detailed geometry cannot be resolved in scale resolving simulations.
These modelled fluctuations can be used as sources in propagation codes or as transient input for Large
Eddy Simulations.

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cooling module. Applied Acoustics. 2016, 111:67-76[4] Lucius A, Hütter S, Dietrich P, Schneider M, Lehmann M, Geyer T. Experimental and numericalinvestigation of axial fan aeroacoustics at disturbed inflow conditions. Proceedings of Fan 2018,
Darmstadt[5] Czwielong F, Krömer F, Becker S. Experimentelle Untersuchung der Schallabstrahlung von
Wärmeübertrager-Axialventilator-Modulen. Proceedings of the DAGA 2019, Rostock[6] Roach PE. The generation of nearly isotropic turbulence by means of grids. Journal of Heat and Fluid
Flow. 1987, 8(2):82-92

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